Vent30th_CVR_C1-C4_Layout 1 12/13/2018 12:09 PM Page 1
INDUSTRIAL VENTILATION
A Manual of Recommended Practice
for Design
INDUSTRIAL
VENTILATION
A Manual of Recommended Practice
for Design
30th Edition
~
30th Edition
ISBN: 978-1-607261-08-7
Defining the Science of
Occupational and Environmental Health®
ACGIH®, Industrial Ventilation: A Manual of Recommended Practice for Design, 30th Edition
Errata Listing (as of 10/9/2019)
CHAPTER
SECTION
PAGE
9
N/A
9-54
11
11.7.6
11-26
DESCRIPTION
Table 9-7 (SI) has been corrected below. Do not use Table 9-7 (SI) as
displayed in the 30th Edition of the ACGIH® publication, Industrial
Ventilation, A Manual of Recommended Practice for Design as it is
incorrect.
Replace last sentence of section with the following: The application
of heat exchangers to industrial exhaust systems is discussed in
Chapter 10 of the ACGIH® Publication, Industrial Ventilation, A
Manual of Recommended Practice for Operation and Maintenance,
which is to be published in 2020.
Table 9-7 (IP). Air Density Correction Factor (Temperature and Elevation Only), df T × dfe
ALTITUDE RELATIVE TO SEA LEVEL (ft)
-5,000 -4,000 -3,000 -2,000 -1,000
0
1,000
2,000
3,000
4,000
5,000
BAROMETRIC PRESSURE
"Hg
35.7
34.5
33.3
32.1
31.0
29.9
28.9
27.8
26.8
25.8
24.9
"wg
485.9
469.5
453.2
436.9
421.9
406.9
393.3
378.4
364.7
351.1
338.9
DENSITY FACTOR (df)
Temp (F)
-40
1.50
1.45
1.40
1.35
1.31
1.26
1.22
1.18
1.13
1.09
1.05
0
1.37
1.32
1.28
1.24
1.19
1.15
1.11
1.07
1.04
1.00
0.96
40
1.26
1.22
1.18
1.14
1.10
1.06
1.02
0.99
0.95
0.92
0.89
70
1.19
1.15
1.11
1.07
1.04
1.00
0.97
0.93
0.90
0.87
0.84
100
1.13
1.09
1.05
1.02
0.98
0.95
0.91
0.88
0.85
0.82
0.79
150
1.03
1.00
0.97
0.93
0.90
0.87
0.84
0.81
0.78
0.75
0.73
200
0.96
0.92
0.89
0.86
0.83
0.80
0.78
0.75
0.72
0.70
0.67
250
0.89
0.86
0.83
0.80
0.77
0.75
0.72
0.70
0.67
0.65
0.62
300
0.83
0.80
0.77
0.75
0.72
0.70
0.67
0.65
0.63
0.60
0.58
350
0.78
0.75
0.73
0.70
0.68
0.65
0.63
0.61
0.59
0.57
0.55
400
0.73
0.71
0.68
0.66
0.64
0.62
0.59
0.57
0.55
0.53
0.51
450
0.69
0.67
0.65
0.62
0.60
0.58
0.56
0.54
0.52
0.50
0.49
500
0.66
0.63
0.61
0.59
0.57
0.55
0.53
0.51
0.50
0.48
0.46
550
0.62
0.60
0.58
0.56
0.54
0.52
0.51
0.49
0.47
0.45
0.44
600
0.60
0.57
0.56
0.54
0.52
0.50
0.48
0.47
0.45
0.43
0.42
700
0.54
0.53
0.51
0.49
0.47
0.46
0.44
0.43
0.41
0.40
0.38
800
0.50
0.48
0.47
0.45
0.44
0.42
0.41
0.39
0.38
0.36
0.35
900
0.46
0.45
0.43
0.42
0.40
0.39
0.38
0.36
0.35
0.34
0.33
1000
0.43
0.42
0.40
0.39
0.38
0.36
0.35
0.34
0.33
0.31
0.30
6,000
7,000
8,000
9,000
10,000
24.0
326.6
23.1
314.4
22.2
302.1
21.4
291.3
20.6
280.4
1.02
0.93
0.85
0.81
0.76
0.70
0.65
0.60
0.56
0.53
0.50
0.47
0.44
0.42
0.40
0.37
0.34
0.31
0.29
0.98
0.89
0.82
0.78
0.73
0.67
0.62
0.58
0.54
0.51
0.48
0.45
0.43
0.41
0.39
0.35
0.33
0.30
0.28
0.94
0.86
0.79
0.75
0.71
0.65
0.60
0.56
0.52
0.49
0.46
0.44
0.41
0.39
0.37
0.34
0.31
0.29
0.27
0.91
0.83
0.76
0.72
0.68
0.63
0.58
0.54
0.50
0.47
0.44
0.42
0.40
0.38
0.36
0.33
0.30
0.28
0.26
0.87
0.80
0.73
0.69
0.66
0.60
0.56
0.52
0.48
0.45
0.43
0.40
0.38
0.36
0.35
0.32
0.29
0.27
0.25
Note that Table 9-7 (SI) has been corrected below. Do not use Table 9-7 (SI) as displayed in
the 30th Edition of the ACGIH book, Industrial Ventilation, A Manual of Recommended
Practice for Design, (i.e., the blue ventilation book) as it is incorrect.
Table 9-7 (SI). Air Density Correction Factor (Temperature and Elevation Only), df T × dfe
ALTITUDE RELATIVE TO SEA LEVEL (m)
-1500
-1200
-900
-600
-300
0
300
600
900
1200
BAROMETRIC PRESSURE
mm Hg
905
875
845
816
787
760
733
707
682
658
kPa
120.7
116.6
112.6
108.7
105.0
101.3
97.8
94.3
91.0
87.7
DENSITY FACTOR (df)
Temp (C)
-10
1.33
1.29
1.24
1.20
1.16
1.12
1.09
1.04
1.01
0.97
0
1.29
1.24
1.20
1.16
1.12
1.08
1.05
1.00
0.97
0.94
10
1.24
1.20
1.15
1.11
1.08
1.04
1.01
0.97
0.94
0.90
20
1.19
1.15
1.11
1.07
1.04
1.00
0.97
0.93
0.90
0.87
30
1.15
1.12
1.08
1.04
1.01
0.97
0.94
0.90
0.87
0.84
40
1.12
1.08
1.04
1.01
0.98
0.94
0.91
0.87
0.85
0.82
50
1.08
1.05
1.01
0.97
0.95
0.91
0.88
0.85
0.82
0.79
60
1.05
1.01
0.98
0.94
0.92
0.88
0.85
0.82
0.79
0.77
70
1.02
0.99
0.95
0.92
0.89
0.86
0.83
0.80
0.77
0.75
80
0.99
0.95
0.92
0.89
0.86
0.83
0.81
0.77
0.75
0.72
90
0.96
0.93
0.90
0.87
0.84
0.81
0.79
0.75
0.73
0.70
100
0.94
0.91
0.88
0.85
0.82
0.79
0.77
0.73
0.71
0.69
120
0.89
0.86
0.83
0.80
0.78
0.75
0.73
0.70
0.68
0.65
140
0.84
0.82
0.79
0.76
0.74
0.71
0.69
0.66
0.64
0.62
160
0.81
0.78
0.75
0.73
0.71
0.68
0.66
0.63
0.61
0.59
180
0.77
0.75
0.72
0.70
0.68
0.65
0.63
0.60
0.59
0.57
200
0.74
0.71
0.69
0.66
0.64
0.62
0.60
0.58
0.56
0.54
250
0.67
0.64
0.62
0.60
0.58
0.56
0.54
0.52
0.50
0.49
300
0.61
0.59
0.57
0.55
0.53
0.51
0.49
0.47
0.46
0.44
400
0.52
0.51
0.49
0.47
0.46
0.44
0.43
0.41
0.40
0.38
500
0.45
0.44
0.42
0.41
0.40
0.38
0.37
0.35
0.34
0.33
600
0.40
0.39
0.38
0.36
0.35
0.34
0.33
0.32
0.31
0.30
700
0.36
0.35
0.33
0.32
0.31
0.30
0.29
0.28
0.27
0.26
1500
1800
2100
2400
2700
3000
634
84.6
611
81.5
589
78.5
567
75.6
546
72.8
526
70.1
0.94
0.91
0.87
0.84
0.81
0.79
0.76
0.74
0.72
0.70
0.68
0.66
0.63
0.60
0.57
0.55
0.52
0.47
0.43
0.37
0.32
0.29
0.25
0.91
0.87
0.84
0.81
0.79
0.76
0.74
0.71
0.70
0.67
0.66
0.64
0.61
0.58
0.55
0.53
0.50
0.45
0.41
0.36
0.31
0.28
0.24
0.87
0.84
0.81
0.78
0.76
0.73
0.71
0.69
0.67
0.65
0.63
0.62
0.59
0.55
0.53
0.51
0.48
0.44
0.40
0.34
0.30
0.27
0.23
0.84
0.81
0.78
0.75
0.73
0.71
0.68
0.66
0.65
0.62
0.61
0.59
0.56
0.53
0.51
0.49
0.47
0.42
0.38
0.33
0.29
0.26
0.23
0.81
0.78
0.75
0.72
0.70
0.68
0.66
0.63
0.62
0.60
0.58
0.57
0.54
0.51
0.49
0.47
0.45
0.40
0.37
0.32
0.27
0.24
0.22
0.78
0.76
0.73
0.70
0.68
0.66
0.64
0.62
0.60
0.58
0.57
0.55
0.53
0.50
0.48
0.46
0.43
0.39
0.36
0.31
0.27
0.24
0.21
INDUSTRIAL
VENTILATION
A Manual of Recommended Practice
for Design
30th Edition
Copyright © 2019
by
ACGIH®
Previous Editions
Copyright © 1951, 1952, 1954, 1956, 1958, 1960, 1962, 1964, 1966, 1968, 1970, 1972, 1974, 1976,
1978, 1980, 1982, 1984, 1986, 1988, 1992, 1995, 1998, 2001, 2004, 2007, 2010, 2013, 2016
by
ACGIH® Industrial Ventilation Committee
_______________________________________________________________
16th Edition — 1980
1st Edition — 1951
17th Edition — 1982
2nd Edition — 1952
18th Edition — 1984
3rd Edition — 1954
19th Edition — 1986
4th Edition — 1956
20th Edition — 1988
5th Edition — 1958
21st Edition — 1992
6th Edition — 1960
22nd Edition — 1995
7th Edition — 1962
23rd Edition — Metric — 1998
8th Edition — 1964
24th Edition — 2001
9th Edition — 1966
25th Edition — 2004
10th Edition — 1968
26th Edition — 2007
11th Edition — 1970
27th Edition — 2010
12th Edition — 1972
28th Edition — 2013
13th Edition — 1974
29th Edition — 2016
14th Edition — 1976
15th Edition — 1978
_______________________________________________________________
ISBN: 978-1-607261-08-7
All rights reserved. Printed in the United States of America. Except as permitted under the United States Copyright Act of
1976, no part of this publication may be reproduced or distributed in any form or by any means or stored in a database or
retrieval system, without prior written permission from the publisher.
ACGIH®
Kemper Woods Center
1330 Kemper Meadow Drive
Cincinnati, Ohio 45240-4148
Telephone: 513-742-2020 Fax: 513-742-3355
Email: Publishing@acgih.org
http://www.acgih.org
CONTENTS
FOREWORD . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .vii
DEDICATION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .ix
ACKNOWLEDGMENTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .x
DEFINITIONS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .xi
ABBREVIATIONS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .xiv
CHAPTER 1
RISK ASSESSMENT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1-1
1.1
Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1-2
1.2
Hazards versus Risks . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1-2
1.3
Risk Assessment . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1-2
1.4
Risk Assessment Process . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1-2
1.5
Airborne Hazard Identification . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1-3
1.6
Exposure Characteristics . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1-7
1.7
Health Hazard Exposure Assessment . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1-9
1.8
Hierarchy of Controls . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1-12
1.9
Summary . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1-13
REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1-15
CHAPTER 2
PRELIMINARY DESIGN AND COST ESTIMATION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-1
2.1
Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-2
2.2
Project Goals and Success Criteria . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-3
2.3
Large Project Team Organization . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-4
2.4
Team Responsibility Matrix (TRM) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-4
2.5
Project Team Safety . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-8
2.6
Document Control . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-8
2.7
Project Team Organization, Selection and Skills . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-8
2.8
Internal Responsibility for Final Approval of Budget, Technical Merit and Regulatory Issues . . . . . . . . . . . .2-9
2.9
Communication of Plant (and Project) Requirements . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-9
2.10 Design/Build, In-House Design or Outside Consultant . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-11
2.11 Design-Construct Method (Separate Responsibilities for Engineering and Installation) . . . . . . . . . . . . . . . .2-14
2.12 Design/Build (Turnkey) Method – Single Source of Responsibility . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-16
2.13 Project Team and System Evaluation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-16
2.14 Project Risk and Non-Performance . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-17
2.15 Using Plant Personnel as Project Resources . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-18
2.16 Interface Between the Plant and Project . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-18
2.17 Impact of New Systems on Plant Operation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-19
2.18 Capital Costs . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-19
2.19 Operating Cost Methods . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-21
2.20 Cost Comparison Methods . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-23
REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-26
CHAPTER 3
PRINCIPLES OF AIRFLOW . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .3-1
3.1
Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .3-2
3.2
Recording Numerical Values . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .3-2
3.3
Properties of Air . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .3-3
3.4
Ideal Gas Law . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .3-5
3.5
Density Factor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .3-6
3.6
Ventilation System Pressures . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .3-7
3.7
Conservation of Mass . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .3-9
3.8
Conservation of Energy . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .3-10
3.9
Psychrometrics . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .3-13
3.10 Dew Points . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .3-21
REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .3-21
iii
iv
Industrial Ventilation
CHAPTER 4
CHAPTER 5
CHAPTER 6
CHAPTER 7
CHAPTER 8
INDUSTRIAL VENTILATION SYSTEM DESIGN PRINCIPLES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4-1
4.1
Administration of Industrial Ventilation System Design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4-2
4.2
Drawings and Specifications . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4-2
4.3
Design Options for Industrial Ventilation Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4-3
4.4
Design Procedures . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4-4
4.5
Distribution of Airflow in Duct Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4-7
4.6
Local Exhaust Ventilation System Types . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4-10
4.7
System Redesign . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4-12
4.8
System Components . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4-12
4.9
Local Exhaust Ventilation System Testing and Balancing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4-15
REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4-15
DUCT SYSTEM AND DISCHARGE STACK DESIGN PRINCIPLES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .5-1
5.1
Duct Systems and Discharge Stacks . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .5-2
5.2
Duct Construction Considerations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .5-2
5.3
Discharge Stacks . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .5-6
REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .5-10
HOOD DESIGN . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-1
6.1
Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-3
6.2
Enclosing Hoods – Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-5
6.3
Totally Enclosing Hoods . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-11
6.4
Hot Processes in Enclosing Hoods . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-13
6.5
Downdraft Occupied Hoods (Clean Rooms) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-13
6.6
Capturing Hoods . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-13
6.7
Choosing Between Capture and Enclosing Hoods . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-28
6.8
Ergonomic Considerations for Design of Hoods . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-28
6.9
Work Practices . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-29
6.10 Material Handling in and Near Hood Workstations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-29
6.11 Hood Maintenance and Cleaning . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-29
6.12 Hoods and Personnel Fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-29
6.13 Ventilation of Radioactive and High Toxicity Processes . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-30
6.14 Determining Hood Static Pressure Losses. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6-31
REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-35
FANS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7-1
Chapter Specific Vocabulary . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7-3
Foreword . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7-4
7.1
Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7-4
7.2
Fan Types . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7-4
7.3
Fan Selection . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7-13
7.4
Fan and System Performance . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7-32
7.5
Fan and System Control . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7-46
7.6
Fan System Effects . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7-54
7.7
Fan Motors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7-71
REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7-73
ACKNOWLEDGMENTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7-73
AIR CLEANING DEVICES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .8-1
8.1
Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .8-2
8.2
Selection of Dust Collection Equipment . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .8-2
8.3
Dust Collector Types . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .8-3
8.4
Additional Aids in Dust Collector Selection . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .8-25
8.5
Control of Mist, Gas and Vapor Contaminants . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .8-25
8.6
Gaseous Contaminant Collectors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .8-25
8.7
Unit Collectors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .8-33
8.8
Dust Collecting Equipment Cost . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .8-33
8.9
Selection of Disposable-Type Air Filtration Equipment . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .8-37
8.10 Radioactive and High Toxicity Operations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .8-39
8.11 Explosion Venting/Deflagration Venting . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .8-40
Contents
CHAPTER 9
CHAPTER 10
CHAPTER 11
REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .8-41
APPENDIX A8 CONVERSION OF POUNDS PER HOUR (EMISSIONS RATE) TO GRAINS
PER DRY STANDARD CUBIC FOOT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .8-42
LOCAL EXHAUST VENTILATION SYSTEM DESIGN CALCULATION PROCEDURES . . . . . . . . . . . . . . . . .9-1
9.1
Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-3
9.2
Preliminary System Design Information . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-4
9.3
Design Considerations for Calculating System Airflow Rates and Resistance Losses . . . . . . . . . . . . . . . . . .9-4
9.4
Static Pressure Losses – Special Considerations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-8
9.5
Basic System Design Procedures and Calculations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-8
9.6
Calculation Sheet Design Procedure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-14
9.7
Sample System Design #1 (Single-Branch System at Standard Air Conditions) . . . . . . . . . . . . . . . . . . . . . .9-16
9.8
Distribution of Airflow in a Multi-Branch Duct System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-22
9.9
Increasing Velocity Through a Junction (Weighted Average Velocity Pressure) . . . . . . . . . . . . . . . . . . . . . .9-24
9.10 System and Fan Pressure Calculations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-25
9.11 The System and Fan Curve Relationship . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-26
9.12 Sample System Design #2 (Multi-Branch System at Standard Air Conditions) . . . . . . . . . . . . . . . . . . . . . . .9-27
9.13 Calculation Methods and Non-Standard Air Density . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-33
9.14 Sample System Design #3 (Single-Branch System at Non-Standard Air Conditions) (IP Units Only) . . . . .9-33
9.15 Sample System Design #4 (Adding a Branch to an Existing System at Non-Standard Air Conditions)
(IP Units Only) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-38
9.16 Air Bleed Design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-41
REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-42
APPENDIX A9 PRESSURE MEASUREMENT IN THE SI SYSTEM . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-42
GENERAL INDUSTRIAL VENTILATION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-1
10.1 Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-3
10.2 Dilution Ventilation Principles . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-4
10.3 Dilution Ventilation for Health . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-4
10.4 Confined Space Ventilation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-11
10.5 Mixtures — Dilution Ventilation for Health . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-14
10.6 Dilution Ventilation for Fire and Explosion (IP Units Only) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-15
10.7 Fire Dilution Ventilation for Mixtures . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-16
10.8 Ventilation for Heat Control . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-16
10.9 Heat Balance and Exchange . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-16
10.10 Acclimatization of the Body . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-17
10.11 Acute Heat Disorders . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-17
10.12 Assessment of Heat Stress and Heat Strain . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-18
10.13 Worker Protection . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-19
10.14 Ventilation Control . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-19
10.15 Ventilation Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-19
10.16 Velocity Cooling . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-22
10.17 Radiant Heat Control . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-22
10.18 Protective Suits for Short Exposures . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-22
10.19 Respiratory Heat Exchangers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-22
10.20 Refrigerated Suits . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-23
10.21 Enclosures . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-23
10.22 Insulation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-23
REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-23
SUPPLY AIR SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .11-1
11.1 Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .11-3
11.2 Purpose of Supply Air Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .11-3
11.3 Supply Air System Design for Industrial Spaces . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .11-7
11.4 Supply Air Equipment . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .11-11
11.5 Supply Air Distribution . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .11-20
11.6 Airflow Rate . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .11-24
11.7 Heating, Cooling and Other Operating Costs . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .11-25
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Industrial Ventilation
11.8 Industrial Exhaust Recirculation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .11-26
11.9 System Control . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .11-30
11.10 System Noise . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .11-31
REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .11-32
CHAPTER 12
SPECIAL TOPICS AND TECHNIQUES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .12-1
12.1 Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .12-2
12.2 Computational Fluid Dynamics . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .12-2
12.3 Combustibility of Dust . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .12-5
12.4 Ventilation Techniques for Engineered Nanomaterials . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .12-7
12.5 EPA Method 204 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .12-13
REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .12-14
CHAPTER 13
SPECIFIC OPERATIONS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .13-1
APPENDICES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .14-1
A Threshold Limit Values for Chemical Substances in the Work
Environment with Intended Changes for 2018 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .14-3
B Physical Constants/Conversion Factors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .14-27
C Testing and Measurement of Ventilation Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .14-35
INDEX . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .15-1
FOREWORD
This 30th edition of ACGIH®’s Industrial Ventilation: A
Manual of Recommended Practice for Design, is to be used
just as the name implies – A Manual of Recommended Practices in the design of industrial ventilation systems. This publication has been developed to serve as a guide to assist in the
control of airborne contaminants that may pose occupational
health hazards to employees. The recommendations provided
herein are intended for use in the practice of industrial hygiene
and are to be interpreted and applied by a person trained in the
discipline. The information contained in this Manual are not
to be construed or used in any way as legal standards, and
ACGIH® does not advocate their use as such. However, it is
recognized that in certain circumstances individuals or organizations may wish to make use of these recommendations as
such. ACGIH® will not oppose use of the Manual in this manner, as in these instances its use will contribute to overall
improvement in worker protection. However, the user must
recognize the constraints and limitations subject to proper use
of the Manual and will bear the responsibility for such use.
Due to the inherent complexity of the science associated
with design of industrial ventilation systems, this edition of the
Manual is written with deference toward a simpler and briefer
means of explanation. Care has been taken to make this manual a practical user’s handbook and not a theoretical treatise.
The reader is encouraged to reference available publications
addressing fluid dynamics should a more comprehensive
understanding of this topic be required.
Special Note to User
This Manual is intended for use in the practice of
industrial hygiene and industrial ventilation design as
guidelines or recommendations to assist in the control of
potential workplace health hazards and for no other use.
These guidelines or recommendations should not be used
by anyone untrained in the discipline of industrial
hygiene or industrial ventilation design. ACGIH® disclaims liability with respect to the use of this Manual.
Practitioners should also note that this Manual contains
techniques and conceptualizations of designs submitted and
adopted through the years as an approach to reduce worker
exposure to airborne contaminants, and as new and better solutions are submitted/approved ACGIH® will update the Manual
with improved design concepts. However, while the techniques in this primer are based on the best available science,
alternative designs may improve upon the conceptualizations
contained herein and are encouraged. If a reader becomes
aware of a better means of protecting workers with the use of
air movement, one is encouraged to submit such a concept to
the ACGIH® Industrial Ventilation Committee for review.
Submissions must be sent to the ACGIH® Science and Education Group by e-mail to science@acgih.org.
Metric (SI) to English (IP) Conversions
Conversion from metric to English is being utilized more
and more in the international, commercial, and regulatory marketplaces. Guidelines have been published and this Manual
uses the U.S. Department of Defense Document SD-10 (published December 2003) for its nomenclature and presentation.
There are some key definitions to be considered going forward in the Manual:
Metric Units (SI)  A system of basic measures defined by
the International Symbol of Units on “Le Système International d’Unités (SI).” These units are described in IEEE/ASTM SI
10.
This Manual has undergone significant changes in the last
few issues and this edition is no exception. Major changes will
be found in many of the chapters herein and conceptual figures
that incorporate computational fluid dynamics (CFD) have
been added for clarification and incorporated for simpler conceptual understanding of air flow patterns.
Inch-Pound Units (IP)  The standards as previously adopted in the United States and some other parts of the world.
There are still some units that have been adopted internationally in some areas.
The chapter addressing General Industrial Ventilation (formerly Chapter 4) has been moved and is now located in Chapter 10 and still includes dilution ventilation techniques and
equations. Chapter 4 now addresses Industrial Ventilation System Design Principles; it provides the overview of local
exhaust ventilation principles. Chapter 5 now addresses Duct
Components and Stack Designs. Finally, Chapter 12 now
addresses Special Topics, including EPA method 204, Computational Fluid Dynamics as a design tool, and combustible dust.
Soft Conversion  The process of changing a measurement
from inch-pound (IP) units to equivalent metric units (SI) with
acceptable measurement tolerances without changing the
physical configuration of the item. For example, one pound =
453.592 grams.
Hard Conversion  The process of changing a measurement
in inch-pound units (IP) to metric units (SI), which necessitates physical configuration changes of the item outside those
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Industrial Ventilation
permitted by established measurement tolerances. For example, one pound = 454 grams.
This Manual uses hard conversions for many of the standards in Chapter 13 and other parts where there are problem
solutions. Soft conversion values are used in tables such as
flash point temperatures and other areas where exact values
were previously published. This will allow for ease of measurement and verification and will provide a more suitable and
comparable solution to problems.
In many cases, where hard conversions provide different
input data (for example, the air volume specified for a particular piece of equipment), problem solutions will obviously
give different results. This should be expected – duct sizes are
different, ranges between duct sizes are different and in many
cases, the SI solution will be more accurate and provide less
energy. This should be weighed against the varied duct sizes
and the problems with fabrication.
INDUSTRIAL VENTILATION COMMITTEE
Jonathan Hale, Air Systems Corporation, North Carolina,
Chair
Gregg Grubb, Grubb Industrial Hygiene Services, LLC,
Michigan, Vice Chair
Lucinette Alvarado Rivera, Covestro, Pennsylvania
Michael Clark, CECO Fisher Klosterman, Tennessee
Robert Dayringer, MIOSHA, Michigan
Frank Demer, Freeport McMoRan Corporation, Arizona
James Friedman, Wood, Inc., Minnesota
Christopher Manning, Materials Processing Solutions, Inc.,
Massachusetts
John McKernan, U.S. EPA, Ohio
Dale Price, M&P Air Components, Inc., California
Rafael Sartim, ArcelorMittal, Brazil
Robert Shearer, KBD Technic, Ohio
Jennifer Topmiller, NIOSH, Ohio
John “Pat” Curran, PCIH, North Carolina – Consultant
Thomas Godbey, Jr., Diagnostic Consultant, Kentucky –
Consultant
Daniel Josephs, Kentucky – Consultant
Gerry Lanham, Ohio (retired) – Consultant
DEDICATION
The ACGIH® Industrial Ventilation Committee dedicates
Industrial Ventilation: A Manual of Recommended Practice for
Design, 30th Edition to our
mentor and colleague, Gerry A.
Lanham, PE. Mr. Lanham is the
former President of KBD/Technic, Inc., and possesses almost
50 years of experience in the
design, testing, and installation
of industrial ventilation systems. He holds a Bachelor of Science degree in Mechanical Engineering from the University
of Cincinnati, and a Master of Business Administration in
Advanced Business Economics degree from Xavier University. Mr. Lanham is the co-author of a patent for HVAC controls and a system for protection against biohazard attack.
ham taught the ACGIH® course, Fundamentals in Industrial
Ventilation & Practical Applications of Useful Equations and
instructed countless other industrial ventilation education
courses. In addition to his service with ACGIH®, Mr. Lanham
served on the following committees: American Society of
Heating, Refrigeration, and Air Conditioning Engineers; the
American Foundry Society; Association of Energy Engineers;
and American National Standards Institute.
Selected as the first recipient of the ACGIH® Robert T.
Hughes Memorial Award (the highest award in industrial ventilation) in 2016, Mr. Lanham was recognized for his contributions to the field of industrial ventilation and ACGIH®. Mr.
Lanham’s outstanding committee leadership and passion for
this Manual led to the metrification of the Manual beginning
with the 28th Edition, which significantly enhanced the usability of the Manual by industrial ventilation practitioners
throughout the world.
Mr. Lanham became a member of ACGIH® in 1996 and
served on the ACGIH® Industrial Ventilation Committee from
that time until 2016; he held the position of Vice Chair from
2008 to 2013 and Chair from 2013 to 2016. First and always
a U.S. Marine, Mr. Lanham lead the ACGIH® Industrial Ventilation Committee by example; he was always focused on
what was best for this Manual, its users, and the health of the
world population. As a member of the Committee, Mr. Lan-
Indeed, the practice of industrial ventilation would be much
less effective without Mr. Lanham’s tireless and egalitarian
efforts. The Committee will surely miss his combination of
integrity, sincerity, and intelligence, and pledges to carry on
with his mission. We offer our best regards to Mr. Lanham as
he spends more time with his family and continues to serve as
a USGA® official. Semper Fi, Gerry!
ix
ACKNOWLEDGMENTS
That dedication has continued for 68 years now and counting. Names below plus numerous supporting consultants and
contributors have donated thousands of hours of their time for
the advancement of the science and art of Industrial Ventilation.
Ken Robinson, one of our founding members, passed away
in June of 2014 at the age of 101. It brought to mind the heritage of our first group that gathered in Lansing, Michigan in
1948 with an idea to publish a Manual that would help industry improve ventilation. Among the names listed below are
names from that early group like Jack Baliff, Ken Morse,
George Hama, Knowlton Kaplan and Norma Donovan. These
were people who not only designed the contents of the Manual
but also personally signed a note to pay for the first publishing
costs.
To all Committee persons and to individuals, companies
and agencies, past and present that have added to the body of
this work we offer our special thanks.
INDUSTRIAL VENTILATION COMMITTEE
Previous Members and Consultants
G.M. Adams, 20042008
L. Alvarado Rivera, 2016–present
A.G. Apol, 19842002
H. Ayer, 19621966
R.E. Bales, 19541960
J. Baliff, 19501956; Chair, 19541956
J.C. Barrett, 1956–1976; Chair, 19601968
J.L. Beltran, 19641966
D. Bonn, Consultant, 19581968
D.J. Burton, 19881990
K.J. Caplan, 19741978; Consultant, 19801986
A.B. Cecala, 19981999
G. Carlton, 19992002
M. Clark, 2016–present
W.M. Cleary, 19762006; Chair, 19781984
J. Curran, Consultant, 2017present
M. Davidson, 19951998
R. Dayringer, 2004present
F.R. Demer, 2016present
L. Dickie, 19841994; Consultant, 19681984
T.N. Do, 19952000
N. Donovan, Editorial Consultant, 19502008
D.L. Edwards, 20032016
B. Feiner, 19561968
M. Flynn, 19891995
M. Franklin, 19911994; 19982001
J.N. Friedman, 2016–present
T. Godbey, Jr., Consultant, 2016present
G. Grubb, 2006present; Vice Chair, 2017present
G.R. Gruetzmacher, 2018–present
S.E. Guffey, 19922015
J.F. Hale, 2004present; Vice Chair, 20132016; Chair,
2017present
G.M. Hama, 19501984; Chair, 19561960
R.L. Herring, 2006–2016
R.P. Hibbard, 19681994
R.T. Hughes, 19872014; Chair, 19892001
G.Q. Johnson, 20012008
H.S. Jordan, 19601962
D. Josephs, Consultant, 2017present
J. Kane, Consultant, 19501952
J. Kayse, Consultant, 19561958
J.F. Keppler, 19501954; 19581960
G.W. Knutson, 19862011
G. Lanham, 19982013; Vice Chair, 20082013; Chair
20142016; Consultant, 2017present
J.J. Loeffler, 19801995; Chair, 19841989
J. Lumsden, 19621968
J.R. Lynch, 19661976
C.P. Manning, 2016present
J.L. McKernan, 2007present
K.R.Mead, 19952001
G. Michaelson, 19581960
K.M. Morse, 19501951; Chair 19501951
J.T. Nalbone, 2016–2017
R.T. Page, 19541956
K.M. Paulson, 19912015; Vice Chair, 19962008
O.P. Petrey, Consultant, 19781999
D. Price, 2016–present
G.S. Rajhans, 19762013; Vice Chair, 19941995; Chair,
20022013
E. Ravert, Consultant, 20152018
K.E. Robinson, 19501954; Chair, 19521954
A. Salazar, 19521954
R. Sartim, 2016–present
E.L. Schall, 19561958
M.M. Schuman, 19621964; Chair, 19681978
R. Shearer, 2016present
J.C. Soet, 19501960
J.L. Topmiller, 2004present
A.L. Twombly, 19872001
J. Willis, Consultant, 19521956
R. Wolle, 19661974
A.W. Woody, 19982015
J.A. Wunderle, 19601964
x
DEFINITIONS
Aerosol: An assemblage of small particles, solid or liquid, suspended in air. The diameter of the particles may vary from
100 microns down to 0.01 micron or less, e.g., dust, fog,
smoke.
Capture Velocity: The air velocity at any point in front of the
hood or at the hood opening necessary to overcome opposing air currents and capture the contaminated air at that
point by causing it to flow into the hood.
Air Cleaner: A device designed for the purpose of removing
atmospheric airborne impurities such as dusts, gases,
mists, vapors, fumes, and smoke. (Air cleaners include air
washers, air filters, electrostatic precipitators, and charcoal filters.)
Comfort Zone (Average): The range of effective temperatures
over which the majority (50% or more) of adults feel comfortable.
Convection: The motion resulting in a fluid from the differences in density and the action of gravity. In heat transmission this meaning has been extended to include both forced
and natural motion or circulation.
Air Filter: An air-cleaning device that removes light particulate loadings from normal atmospheric air before introduction into the building. Usual range: loadings up to 3 grains
per thousand cubic feet (0.003 grains per cubic foot
[0.00687 grams/m3]). Note: Atmospheric air in heavy industrial areas and in-plant air in many industries have higher
loadings than this, and dust collectors are then indicated for
proper air cleaning.
Deflagration: A propagation of a combustion zone that occurs
at a velocity that is less than the speed of sound in the unreacted medium.
Density: The ratio of the mass of a specimen of a substance to
the volume of the specimen. The mass of a unit volume of
a substance. When weight can be used without confusion, as
synonymous with mass, density is the weight of a unit volume of a substance.
Air Horsepower: The theoretical horsepower required to drive
a fan if there were no losses in the fan; that is, if its efficiency were 100 percent.
Density Factor: The ratio of actual air density to density of standard air. The product of the density factor and the density of
standard air (0.075 lb/ft3[1.204 kg/m3]) will give the actual air
3
density in pounds per cubic foot; Density = df H 0.075 lb/ft
(the density of standard air).
Aspect Ratio: The ratio of the width to the length; AR = W/L.
Aspect Ratio of an Elbow: The width (W) along the axis of the
bend divided by depth (D) in the plane of the bend; AR =
W/D.
Balanced Industrial Ventilation System: Installed and reliably
operating on a continuous basis and meets following criteria: 1) Airflows are a minimum at all hoods to meet capture
of pollutants to protect operator and plant environment, 2)
Transport (Conveying) Velocity is maintained in all branches and main lines carrying air, 3) System is operating at
minimum System Static Pressure and Power as designed, 4)
Fan is operating at stable point on fan curve and is properly
controlled.
Dust: Small solid particles created by the breaking up of larger
particles by processes, i.e., crushing, grinding, drilling,
explosions, etc. Dust particles already in existence in a mixture of materials may escape into the air through such operations as shoveling, conveying, screening, sweeping, etc.
Blast Gate: Sliding damper.
Entry Loss: Loss in pressure caused by air flowing into a duct
or hood.
Dust Collector: An air-cleaning device to remove heavy particulate loadings from exhaust systems. Usual range of particulate loading: 0.003 grains per cubic foot [0.00687 grams/m3]
or higher.
Blow (throw): In air distribution, the distance an air stream
travels from an outlet to a position at which air motion
along the axis reduces to a velocity of 50 fpm [0.254 m/s].
For unit heaters, the distance an air stream travels from a
heater without a perceptible rise due to temperature difference and loss of velocity.
Fumes: Small, solid particles formed by the condensation of
vapors of solid materials.
Gases: Formless fluids that tend to occupy an entire space uniformly at ordinary temperatures and pressures.
Hard Conversion: In a hard conversion a new rationalized
number is created that is convenient to work with and
remember. Example 48 × 25 = 1200 mm
Brake Horsepower: The horsepower actually required to drive
a fan. This includes the energy losses in the fan and can be
determined only by actual test of the fan. (This does not
include the drive losses between motor and fan.)
Hood: A shaped inlet designed to capture contaminated air and
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Industrial Ventilation
conduct it into the exhaust duct system.
Hood Flow Coefficient: The ratio of flow caused by a given
hood static pressure compared to the theoretical flow that
would result if the static pressure could be converted to
velocity pressure with 100 percent efficiency. NOTE: This
was defined as Coefficient of Entry in previous editons.
Humidity, Absolute: The weight of water vapor per unit volume (pounds per cubic foot or grams per cubic centimeter).
Humidity, Relative: The ratio of the actual partial pressure of
the water vapor in a space to the saturation pressure of pure
water at the same temperature.
Inch of Water: A unit of pressure equal to the pressure exerted by a
column of liquid water one inch high at a standard temperature.
Lower Explosive Limit: The lower limit of flammability or
explosibility of a gas or vapor at ordinary ambient temperatures expressed in percent of the gas or vapor in air by volume.
This limit is assumed constant for temperatures up to 250 F
[120 C]. Above these temperatures, it should be decreased by
a factor of 0.7 since explosibility increases with higher temperatures.
Manometer: An instrument for measuring pressure; essentially
a U-tube partially filled with a liquid, usually water, mercury or a light oil, so constructed that the amount of displacement of the liquid indicates the pressure being exerted
on the instrument.
Micron: A unit of length, the thousandth part of 1 mm or the
millionth of a meter (approximately 1/25,000 of an inch).
Minimum Design Duct Velocity: Minimum air velocity
required to move the particulates in the air stream (fpm
[m/s]).
Mists: Small droplets of materials that are ordinarily liquid at
normal temperature and pressure.
Plenum: Pressure equalizing chamber.
Plug Flow: The velocity of a fluid is assumed to be constant
across any cross section of the duct perpendicular to the axis
of the duct.
Pressure, Static: The potential pressure exerted in all directions by a fluid at rest. For a fluid in motion, it is measured
in a direction normal to the direction of flow. Usually
expressed in inches water gauge when dealing with air. (The
tendency to either burst or collapse the pipe.)
Pressure, Total: The algebraic sum of the velocity pressure and
the static pressure (with due regard to sign).
Pressure, Vapor: The pressure exerted by a vapor. If a vapor is
kept in confinement over its liquid so that the vapor can
accumulate above the liquid, the temperature being held con-
stant, the vapor pressure approaches a fixed limit called the
maximum or saturated vapor pressure, dependent only on
the temperature and the liquid. The term vapor pressure is
sometimes used as synonymous with saturated vapor pressure.
Pressure, Velocity: The kinetic pressure in the direction of flow
necessary to cause a fluid at rest to flow at a given velocity.
Usually expressed in inches water gauge.
Radiation, Thermal (Heat): The transmission of energy by
means of electromagnetic waves of very long wavelength.
Radiant energy of any wavelength may, when absorbed,
become thermal energy and result in an increase in the temperature of the absorbing body.
Replacement Air: A ventilation term used to indicate the volume of controlled outdoor air supplied to a building to
replace air being exhausted.
Slot Velocity: Linear flow rate of contaminated air through a
slot, fpm.
Smoke: An air suspension (aerosol) of particles, usually but not
necessarily solid, often originating in a solid nucleus,
formed from combustion or sublimation.
Soft Conversion: An inch-pound measurement or value is converted to its exact or nearly exact metric equivalent. Example 48 × 25.4 = 1219.4 mm
Specific Gravity: The ratio of the mass of a unit volume of a
substance to the mass of the same volume of a standard substance at a standard temperature. Water at 39.2 F [4 C] is the
standard substance usually referred to. For gases, dry air, at
the same temperature and pressure as the gas, is often taken
as the standard substance.
Standard Air: Dry air at 70 F and 29.92 (in Hg) barometer.
This is substantially equivalent to 0.075 lb/ft3. Specific heat
of dry air = 0.24 BTU/lb/F. For this Manual, the equivalent
values in SI system are: Density = 1.204 kg/m3 @ 21 C @
Sea Level. CP = 1.005 kJ/kgK.
Temperature, Effective: An arbitrary index that combines into a
single value the effect of temperature, humidity, and air movement on the sensation of warmth or cold felt by the human
body. The numerical value is that of the temperature of still,
saturated air that would induce an identical sensation.
Temperature, Wet-Bulb: Thermodynamic wet-bulb temperature
is the temperature at which liquid or solid water, by evaporating into air, can bring the air to saturation adiabatically at
the same temperature. Wet-bulb temperature (without qualification) is the temperature indicated by a wet-bulb psychrometer constructed and used according to specifications.
Threshold Limit Values (TLVs®): The values for airborne toxic
General Industrial Ventilation
materials that are to be used as guides in the control of
health hazards and represent time-weighted concentrations
to which nearly all workers may be exposed for 8 hours per
day over extended periods of time without adverse effects
(see Appendix).
Transport (Conveying) Velocity: See Minimum Design Duct
Velocity.
Turn-Down Ratio: The degree to which the operating performance of a system can be reduced to satisfy part-load conditions. Usually expressed as a ratio; for example, 30:1
means the minimum operation point is 1/30th of full load.
xiii
Vapor: The gaseous form of substances that are normally in
the solid or liquid state and that can be changed to those
states either by increasing the pressure or decreasing the
temperature.
ABBREVIATIONS
A . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .area
AC . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .alternating current
acfm . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .actual flow rate
Af . . . . . . . . . . . . . . . . . . . . . . . . . .face area of hood opening
AHP . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .air horsepower
am3/s . . . . . . . . . . . . . . . . . . . . . . . . . .actual cubic meters/sec
AR . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .aspect ratio
As . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .slot area
ASL . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .above sea level
B . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .barometric pressure
BHP . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .brake horsepower
BHPa . . . . . . . . . . . . . . . . . . . . . . . .brake horsepower, actual
BHPs . . . . . . . . . . . . . . . . . . . .brake horsepower, standard air
BTU . . . . . . . . . . . . . . . . . . . . . . . . . . . .British Thermal Unit
BTUh . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .BTU per hour
C . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .degrees, Celsius
calc sheet . . . . . . . . . . . . . . . . . . . .ACGIH® calculation sheet
Ccap . . . . . . . . . . . . . . . . . . . . .hood configuration flow factor
Ce . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .hood flow coefficient
CLR . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .centerline radius
Cp . . . . . . . . . . . . . . .heat capacity in BTU/lbm-R [kJ/kg K]
d . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .diameter
D . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .depth
da . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .dry air
dB . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .decibels
dBA . . . . . . . . . . . . . . . . .“A” weighted sound pressure level
DC . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .direct current
dequiv . . . . . . . . . . . . . . . . . . . . . . . . . . . . .equivalent diameter
df . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .density factor
dfe . . . . . . . . . . . . . . . . . . . . . . . . . . . .elevation density factor
dfm . . . . . . . . . . . . . . . . . . . . . . . . . . . .moisture density factor
dfp . . . . . . . . . . . . . . . . . . . . . . . . . . . .pressure density factor
dft . . . . . . . . . . . . . . . . . . . . . . . . . .temperature density factor
dnm3/s . . . . . . . . . . . . . .dry normal cubic meters per second
dscf . . . . . . . . . . . . . . . . . . . . . . . . . . .dry standard cubic feet
dscfm . . . . . . . . . . . . . . . .dry standard cubic feet per minute
ET . . . . . . . . . . . . . . . . . . . . . . . . . . . . .effective temperature
f . . . . . . . . . . . . . . . . . . . .Moody diagram friction coefficient
F . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .degrees, Fahrenheit
Fa . . . . . . . . . . . . .acceleration (or Bernoulli) coefficient = 1
Fcont . . . . . . . . . . . . . . . . . . . . . . . . . . . .contraction loss factor
FCXP . . . . . . . . . . . . . . . . . . . . . .fan cooled explosion proof
Fd . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .duct loss factor
F'd . . . . . . . . . . . . . . . . . . . . . . . . . .loss per unit length (duct)
Fel . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .elbow loss factor
Fen . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .entry loss factor
Fexp . . . . . . . . . . . . . . . . . . . . . . . . . . .expansion regain factor
Fh . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .duct entry loss factor
FLA . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .full load amps
fpm . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .feet per minute
fps . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .feet per second
Fs . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .slot loss factor
FSP . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .fan static pressure
FTP . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .fan total pressure
ft . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .feet
ft2 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .square feet
ft3 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .cubic feet
g . . . . . . . . . . . . . . . . . . . . . . . . .gravitational force, ft/sec/sec
gpm . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .gallons per minute
gr . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .grains
h . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .total heat (enthalpy)
H . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .height
HEPA . . . . . . . . . . . . . . . . . . . .high-efficiency particulate air
hp . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .horsepower
hr . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .hour
Hslot . . . . . . . . . . . . . . . . . . . . . . . . .height of hood, table, slot
HV . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .humid volume
HVAC . . . . . . . . . .heating, ventilation, and air conditioning
in . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .inches
in2 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .square inches
IVS . . . . . . . . . . . . . . . . . . . . . . .industrial ventilation system
K . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .degrees, Kelvin
kg . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .kilograms
kPa . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .kilo Pascals
L . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .length
lb . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .pound
lbf . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .pounds force
lbm . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .pounds mass
LEL . . . . . . . . . . . . . . . . . . . . . . . . . . . .lower explosive limit
LEV . . . . . . . . . . . . . . . . . . . . . . . . .local exhaust ventilation
LP . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .sound pressure level
Lw . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .sound power level
m . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .meters
ṁ . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .mass flow rate
M . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .molar weight
m2 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .square meters
m3 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .cubic meters
ME . . . . . . . . . . . . . . . . . . . . . . . . . . . .mechanical efficiency
mg . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .milligram
min . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .minutes
xiv
General Industrial Ventilation
mm . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .millimeters
mm wg . . . . . . . . . . . . . . . . . . . . . . .millimeters water gauge
MRT . . . . . . . . . . . . . . . . . . . . . . . .mean radiant temperature
m/s . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .meters per sec
MW . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .molecular weight
n . . . . . . . . . . . . . . . . . . . . . . . . . . .number of moles; normal
N . . . . . . . . . . . . . . . . . . . . . . . . . . . . .rotational speed (RPM)
h . . . . . . . . . . . . . . . . . . . . . . . . . . . .efficiency, fan efficiency
hn . . . . . . . . . . . . . . . . . . . . . . . . . . . . .humidifying efficiency
hs . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .fan static efficiency
hT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .fan total efficiency
NX . . . . . . . . . . . . . . . . . . . . . . . . . . .synchronous speed, rpm
ODP . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .open drip proof
p . . . . . . . . . . . . . . . . . . . . . . . . . . . .number of poles (motor)
P . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .pressure
r . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .density of air
Pa . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .Pascals
Pa . . . . . . . . . . . . . . . . . . . . . . . . . .actual pressure in wg [Pa]
Pequiv . . . . . . . . . . . . . . . . . . . . . . . . . . . . .equivalent pressure
PF . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .power factor
ppm . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .parts per million
psi . . . . . . . . . . . . . . . . . . . . . . . . . . . .pounds per square inch
PWR . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .power
Q . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .flow rate
Qact . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .actual airflow rate
Qcorr . . . . . . . . . . . . . . . . . . . . . . . . . . . .corrected airflow rate
Qstd . . . . . . . . . . . . . . . . . . . . . . . . . . . . .standard airflow rate
r . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .radius
R . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .degrees, Rankin
Rg . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .specific gas constant
RH . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .relative humidity
RPM . . . . . . . . . . . . . . . . . . . . . . . . . .revolutions per minute
s . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .seconds
scfm . . . . . . . . . . . . . . . . . . . .standard cubic feet per minute
SEF . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .system effect factor
SEL . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .system effect loss
sfpm . . . . . . . . . . . . . . . . . . . . . . . . . .surface feet per minute
SG . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .specific gravity
xv
SPf . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .filter pressure loss
SPgov . . . . . . . . . . . . . . . . . . . . . . . .governing static pressure
SPh . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .hood static pressure
SPL . . . . . . . . . . . . . . . . . . . . . . . . . . . . .sound pressure level
SPs . . . . . . . . . . . . . . . . . . . .SP, system handling standard air
SSP . . . . . . . . . . . . . . . . . . . . . . . . . . . .system static pressure
STP . . . . . . . . . . . . . . . . . .standard temperature and pressure
SWL . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .sound power level
t . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .time
T . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .temperature
TEFC . . . . . . . . . . . . . . . . . . . . . .totally enclosed fan cooled
TLV® . . . . . . . . . . . . . . . . . . . . . . . . . .Threshold Limit Value
TP . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .total pressure
V . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .velocity
Vd . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .duct velocity
Vf . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .hood face velocity
VFD . . . . . . . . . . . . . . . . . . . . . . . . .variable frequency drive
VIV . . . . . . . . . . . . . . . . . . . . . . .variable inlet vane (damper)
VP . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .velocity pressure
VPd . . . . . . . . . . . . . . . . . . . . . . . . . . . .duct velocity pressure
VPin . . . . . . . . . . . . . . . . . . . . . .velocity pressure at fan inlet
VPout . . . . . . . . . . . . . . . . . . . . .velocity pressure at fan outlet
VPr . . . . . . . . . . . . . . . . .weighted-average velocity pressure
VPs . . . . . . . . . . . . . . . . . . .slot or opening velocity pressure
Vs . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .slot velocity
Vt . . . . . . . . . . . . . . . . . . . . . . . . . . . . .duct transport velocity
VX . . . . . . . . . . . . . .capture velocity necessary at distance X
from the hood face
w . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .watt
W . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .width
ẇ . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .work
w . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .moisture content
Wfl . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .flange width
"wg . . . . . . .inches water gauge (pressure unit in IP system)
WK2 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .inertia
X . . . .greatest distance between contaminant and hood face
z . . . . . . . . . . . . . . . . . . . . . . . . . . . .elevation above sea level
Chapter 1
RISK ASSESSMENT
NOTE: Equations with notation followed by (IP) are designated for inch-pound system only; equations followed by (SI) are designated for
metric use only. If equation bears neither, then it applies to both systems.
1.1
1.2
1.3
1.4
1.5
1.6
INTRODUCTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1-2
HAZARDS VERSUS RISKS . . . . . . . . . . . . . . . . . . . . .1-2
RISK ASSESSMENT . . . . . . . . . . . . . . . . . . . . . . . . . . .1-2
RISK ASSESSMENT PROCESS . . . . . . . . . . . . . . . . . .1-2
1.4.1 Anticipate, Identify and Analyze the Hazards . .1-3
1.4.2 Assess the Exposures . . . . . . . . . . . . . . . . . . . . .1-3
1.4.3 Estimate the Risks . . . . . . . . . . . . . . . . . . . . . . .1-3
1.4.4 Determine the Appropriate Controls . . . . . . . . .1-3
1.4.5 Record Findings . . . . . . . . . . . . . . . . . . . . . . . . .1-3
1.4.6 Monitor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1-3
AIRBORNE HAZARD IDENTIFICATION . . . . . . . . .1-3
1.5.1 Physical Forms . . . . . . . . . . . . . . . . . . . . . . . . . .1-3
1.5.2 Mechanisms of Generation and Dispersion . . . .1-3
1.5.3 Hazard Types . . . . . . . . . . . . . . . . . . . . . . . . . . .1-7
1.5.4 Hazard Degrees . . . . . . . . . . . . . . . . . . . . . . . . .1-7
EXPOSURE CHARACTERISTICS . . . . . . . . . . . . . . . .1-7
1.6.1 Amount, Duration and Frequency . . . . . . . . . . .1-8
1.6.2 Conditions of Generation and Dispersion . . . . .1-8
1.6.3 Physical Form . . . . . . . . . . . . . . . . . . . . . . . . . . .1-9
1.6.4 Volatility . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1-9
1.6.5 Aerodynamic Diameter . . . . . . . . . . . . . . . . . . .1-9
1.6.6 Moisture Content . . . . . . . . . . . . . . . . . . . . . . . .1-9
1.7 HEALTH HAZARD EXPOSURE ASSESSMENT . . . .1-9
1.7.1 Observations and Monitoring . . . . . . . . . . . . .1-10
1.7.2 Surrogate Data . . . . . . . . . . . . . . . . . . . . . . . . .1-11
1.7.3 Modeling . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1-11
1.7.4 Occupational Exposure Limits . . . . . . . . . . . . .1-11
1.7.5 Assessing Exposures . . . . . . . . . . . . . . . . . . . .1-12
1.8 HIERARCHY OF CONTROLS . . . . . . . . . . . . . . . . . .1-12
1.8.1 Elimination or Substitution . . . . . . . . . . . . . . .1-12
1.8.2 Engineering Controls . . . . . . . . . . . . . . . . . . . .1-12
1.8.3 Administrative Controls . . . . . . . . . . . . . . . . . .1-13
1.8 4 Personal Protective Equipment . . . . . . . . . . . .1-13
1.9 SUMMARY . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1-13
REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1-15
____________________________________________________________
Figure 1-1
Figure 1-2
Figure 1-3
Figure 1-4
Evaporation of Volatile Liquid . . . . . . . . . . . . . .1-8
Dust Expulsion by Mechanical Compression . .1-8
Dust Generated from Falling Materials . . . . . . .1-8
Displaced Air Containing Fine Particulate . . . .1-8
Figure 1-5
Figure 1-6
Figure 1-7
Dust Created by Abrasive Blasting . . . . . . . . . .1-9
Hierarchy of Exposure Control Measures . . . .1-13
Summary of Risk Assessment Process with
Emphasis on Airborne Hazards . . . . . . . . . . . .1-14
____________________________________________________________
Table 1-1
Table 1-2
Table 1-3
Table 1-4
Example Risk Assessment Matrix . . . . . . . . . .1-4
Example Risk Assessment Documentation
Form with Example Risk Assessment . . . . . . . .1-5
Generation and Dispersion Mechanisms for
Airborne Hazards . . . . . . . . . . . . . . . . . . . . . . . .1-6
Industrial Processes Known to Produce
Airborne Hazards . . . . . . . . . . . . . . . . . . . . . . . .1-7
Table 1-5
Table 1-6
Table 1-7
Table 1-8
Types of Aerosols . . . . . . . . . . . . . . . . . . . . . . .1-7
Airborne Hazard Categories and Types . . . . . .1-10
Likelihood of Exposure Indicators for
Airborne Hazards . . . . . . . . . . . . . . . . . . . . . . .1-11
Commonly-used Formal Occupational
Exposure Limits in the United States . . . . . . .1-12
1-2
Industrial Ventilation
1.1
INTRODUCTION
Ventilation controls are one means of improving environmental conditions and safety in the workplace by controlling
risks that may be associated with airborne hazard. As with any
risk management endeavor, the design of ventilation controls
should always be preceded by, and be based on, a thorough
understanding of the problem. Problem characterization is
essential in order to anticipate and recognize potential hazardous exposures, evaluate and prioritize the risk, and formulate effective controls. This chapter introduces the basic risk
assessment process as a framework and tool for understanding
airborne hazards that present risks that can be managed by various methods, including ventilation controls.
1.2
HAZARDS VERSUS RISKS(1.2)
Before introducing the basic risk assessment process, it is
important to understand the difference and relationship
between the terms “Hazard” and “Risk.” Hazard is the source
of harm. Hazards cause different types of harm and come in
varying degrees. Hazards include the characteristics of things
(e.g., chemical, biological or radiological agents, equipment,
technology, processes) and the actions or inactions of people.
Risk, on the other hand, is the chance that there could be harm
from a hazard, together with an indication of how serious the
harm could be. Depending on the likelihood of exposure to a
particular hazard or hazards, risks could be the chance of:
varying degrees of acute or chronic health effects or injuries to
people; environmental damage; equipment damage; business
interruption; legal liability, or other harms of concern.
Hazard is just the source of harm and actual harm only
occurs when there is exposure. Risk is the likelihood of actual
harm. It is a function of the hazard and the likelihood of exposure. Simply stated – “Risk” can be considered a product of the
“Hazard and the severity of its potential harm” and the
“Likelihood of exposure” (Equation 1.1).
Risk = (Hazard and the severity of its potential harm)
H (Likelihood of exposure)
1.3
[1.1]
RISK ASSESSMENT
A risk assessment is basically a careful examination or analysis of the hazards and the likelihood of exposure, in order to
determine if enough precautions have been taken or if more
should be done to prevent harm. The goal of a risk assessment
is to determine appropriate ways to eliminate or control hazards and minimize risks.
1.4
(1.2)
RISK ASSESSMENT PROCESS
While the control of airborne hazards is the focus of this
Manual, it is important to approach each problem as a whole.
When considering ventilation controls to address an unacceptable risk from airborne hazards, all hazards and risks need to
be assessed. This could include hazardous energy, pinch
points, working at elevated heights, ergonomic hazards, noise,
heat stress, etc. All of these may influence the application of
ventilation controls and should be considered along with any
airborne hazards in the assessment. Failing to do so could
inadvertently create additional hazards and/or risks, or could
miss an opportunity to concurrently address existing hazards
and/or risks. For example, if an ergonomic hazard for a job,
such as sustained, poor worker posture, is not considered in a
ventilation design, the workers may make modifications to the
design to improve ergonomic conditions and defeat the effective function of the ventilation system.
Risk assessments should be proactive and conducted as
early as possible – ideally in the work design phase. The work
design phase is where it is easiest to most efficiently and effectively minimize hazards and risk.
Risk assessments should be done by a multi-disciplined and
experienced team of individuals who have knowledge of the
work and workplace, its hazards and their controls. The size
and make-up of the team will vary depending on the complexity of the problem. The team should include worker representation, as the workers are often most knowledgeable about the
work and workplace. Other subject matter experts, including
industrial hygienists, safety professionals, environmental professionals, process engineers, mechanical engineers, and legal
experts, should be included as needed. A qualified industrial
hygienist should be included whenever hazards are involved
that could adversely affect the health of workers or the environment. This is particularly important when the hazards are
airborne as industrial ventilation is a core competency in the
industrial hygiene profession.
The risk assessment should be limited to a manageable process, material, system or equipment. Typically, the work is
broken down to the task level and an assessment is performed
on each task. All aspects of the work should be assessed
including:
1) Routine and non-routine activities (e.g., construction,
maintenance, repair and cleaning);
2) Failure modes, including historical and foreseeable
uses and misuses of facilities, materials, and equipment,
3) Existing controls, and
4) All people and things with a likelihood of exposure
(e.g., workers, contractors, public, surrounding environment, etc.).
Information on the workplace, workforce and hazards
should be gathered to perform the assessment. Various sources
of information should be reviewed and could, for example,
include:
1) Inspection of the workplace
2) Interviews of current and intended system users
3) Review of standard operating procedures
4) Task observations
Risk Assessment
5) Identification and review of applicable codes, regulations, internal standards and consensus standards
6) Review of Safety Data Sheets (SDSs)
7) Review of system specifications
8) Review of historical information and data (e.g., industry experience, exposure data, accident records, incident investigations reports, and government, regulatory
agency and manufacturer’s literature)
There is no single recommended method for conducting a
risk assessment. Various standards and guidance documents
exist and should be referred to for more details. All involve, at
least, the following basic steps that are outlined below.
1.4.1 Anticipate, Identify and Analyze the Hazards. This
is one of the most important steps in conducting a risk assessment. An unrecognized hazard cannot be controlled. All significant hazards should be considered. Airborne hazards and their
evaluation are described in Section 1.5.
1.4.2 Assess the Exposures. The potential for exposure
may be obvious based on knowledge and experience. In other
cases, more information may need to be gathered. In each
case, who or what might be harmed must be established and
the likelihood of exposure determined or estimated. Factors
that affect the likelihood of exposure for airborne hazards and
assessing health hazard exposures are discussed in Sections
1.6 and 1.7.
1.4.3 Estimate the Risks. All available hazard and exposure information should be considered along with professional
judgement, to determine the best estimate of risk. A risk
assessment matrix, based on the severity of harm and the likelihood of exposure, can be used to standardize risk determinations. Table 1-1 provides an example of a basic risk assessment
matrix. Such a matrix can be customized to address the organization’s particular risks of concern (e.g., acute or chronic
health effects or injuries to people; environmental damage;
equipment damage; business interruption; legal liability, etc.)
and risk tolerance. The matrix can also be used to summarize
and prioritize or rank risks for communication purposes and to
aid in resource allocation decision making.
1.4.4 Determine the Appropriate Controls. Risks that are
deemed high or serious should be controlled. Those that are
considered medium may require additional assessment and/or
monitoring. Risks that are judged to be low do not require any
immediate action. The hierarchy of controls (Section 1.8)
should be used to eliminate or reduce the hazards and/or
reduce the likelihood of exposure and minimize the risks.
Industrial ventilation, an engineering control, is just one of
many control method options in the hierarchy.
1.4.5 Record Findings. The assessment process should be
documented and include a risk estimate for pre and post additional controls. The risks still present after additional controls
are termed “residual risks.” The findings can be used to communicate and manage the risks. Table 1-2 provides a sample
1-3
records form and work task to show how the form could be
used. It is based on a risk assessment that preceded the development of the ventilation design shown in Chapter 13, VS- 9909.
1.4.6 Monitor. The process should be performed pre and
post control implementation to assess the actual effectiveness
of the controls and determine if additional controls are necessary based on the residual risks. Monitoring should continue
until an acceptable risk level is obtained and periodically conducted to ensure the risk remains tolerable.
1.5
AIRBORNE HAZARD IDENTIFICATION
Industrial ventilation is used to control airborne hazard
exposures. There are a variety of ventilation design criteria and
control techniques available. Their selection will depend primarily on the physical form and the type and degree of the hazard.
1.5.1 Physical Forms. Airborne hazards can be in the physical form of vapors, gases or aerosols. Physical forms behave
differently in air and may require distinct ventilation control
strategies. Vapors are the gaseous state of substances that are a
liquid or solid at Normal Temperature and Pressure (68 F or
20 C and 760 mmHg).(1.7) Evaporation is the process by which
a liquid changes to the vapor state. Sublimation is the process
by which a solid changes directly to the vapor state. Gases are
any substance that is in a gaseous state at Normal Temperature
and Pressure.(1.7) Aerosols are solid or liquid particles suspended in air and include mists, dust, fibers, smokes, fogs and
fumes (Table 1-4).
1.5.2 Mechanisms of Generation and Dispersion. Energy
is required to generate and disperse airborne hazards. The form
of energy could be thermal, mechanical and/or chemical.
Often, more than one energy is involved in a process. Table 12 describes the types of mechanisms and processes known to
generate airborne hazards and their potential physical form(s).
Understanding how airborne hazards are generated aids in
understanding when an industrial ventilation system is
required. Certain industrial processes, such as welding, spray
painting and abrasive blasting, are known to produce airborne
hazards and often, ventilation is the most appropriate hazard
control (Table 1-5).
Outgassing, evaporation (Figure 1-1) and sublimation are
thermal mechanisms by which gases, liquids and some solids
become airborne. Increasing heat increases the generation of
airborne hazards by this mechanism. Heating may also result
in combustion or oxidation of the material (a chemical reaction) and this changes the physical form and chemical composition of the airborne hazard to a gas or smoke. In addition,
heat can transport airborne hazards upwards by convection.
Cooling of heated vapors can result in condensation, changing
in the physical form of the hazard to a fog or fume.
Increasing motion is a mechanical means of creating airborne hazards in liquids and solid materials, creating mists,
1-4
Industrial Ventilation
TABLE 1-1. Example Risk Assessment Matrix (adapted from Department of Defense MIL-STD-882E Standard Practice for System Safety) (1.1)
TABLE 1-2. Example Risk Assessment Documentation Form with Example Risk Assessment
Risk Assessment
1-5
1-6
Industrial Ventilation
TABLE 1-3. Generation and Dispersion Mechanisms for Airborne Hazards
dusts or fibers. Examples of mechanical processes include
vibrating, mixing, splashing, compressing (Figure 1-2) and
sweeping. Increasing motion generally increases the generation and/or dispersion of airborne hazards by this mechanism.
Mechanical breakdown produces airborne hazards by
reducing the source materials to a size that can become airborne. The motion also disperses airborne hazards. The dispersion is often directional and can be at significant velocity.
Scraping, breaking, excavating, drilling, machining, grinding,
sanding, cutting and crushing are some example mechanical
breakdown processes. The rate of mechanical breakdown
increases with the energy associated with the process. Friable
materials are more prone to mechanical breakdown.
Free falling liquids and fine solid materials (i.e., dusts,
fibers) can create airborne hazards (Figure 1-3). This mechanism not only releases airborne hazards, it also generates
smaller aerosols from larger ones by impaction and friction.
The sudden compaction of a falling mass of material can expel
airborne hazards at impact. In addition, air can be entrained in
the falling mass and strip airborne hazards from the material as
it falls (pressure differentials). The falling height is an important influence on the generation and dispersion of airborne
hazards by this mechanism.
Pressure differentials can cause fluid movement in the form
of gas and/or liquids. Bubbling, displacement (Figure 1-4),
blowing, spraying, blasting (Figure 1-5), pressure cutting and
exploding are some example processes where this takes place.
As with mechanical breakdown, the gas and/or fluid movement can cause directional discharge of airborne hazards and
the velocity of the discharge can be significant.
Lastly, certain chemical reactions can generate airborne
hazards. The resulting materials will have a different chemical
composition than the source materials or reactants. They may
also have different forms than the source materials and may
have completely different hazard types and hazard degrees
(Section 1.5.3 and 1.5.4). The rate of chemical reactions
Risk Assessment
TABLE 1-4. Industrial Processes Known to Produce Airborne
Hazards
depends on many factors including: form, concentration, order
of reactants, temperature, pressure, catalysts, solvents, mixing,
etc.
1.5.3 Hazard Types. Airborne substances may be health
hazards, which could be chemical, biological or radiological in
nature. Exposure may cause acute or chronic health effects.
Inhalation is the primary route of exposure in the workplace
and often the primary reason for industrial ventilation systems.
Airborne substances may also be physical hazards such as
flammable or combustible, reactive or explosive and industrial
ventilation systems are often used to control these hazards
also. Physical hazards may manifest as fires, explosions,
excessive temperatures, or the release of large volumes of gas
or toxic or flammable gases or vapors. Industrial ventilation
systems can also be used to control air quality conditions, such
as temperature and humidity. These conditions can not only
affect comfort and cause health effects in individuals, but can
TABLE 1-5. Types of Aerosols
1-7
cause physical damage as well. Table 1-6 provides some
examples of the airborne hazards by category and type.
1.5.4 Hazard Degrees. Both health and physical hazards
come in varying degrees and each hazard category can have
difference potentials for causing harm. The hazard type and
category determine the severity of potential harm. Various
physical, chemical and toxicological characteristics are used to
classify the degree of hazard. For instance, average lethal dose
(LD50) can be used to differentiate various hazard degrees of
acute toxicity and flash point, and flammable range are often
used to distinguish flammability hazard degrees. The United
Nations’ Globally Harmonized System of Classification and
Labeling of Chemicals (GHS) defines various hazard degrees
for chemical hazards.(1.6) Biological and radiological hazards
may have different hazard degrees, as well, based on characteristics such as infectious dose or radiation activity. Thermal
stress and comfort issues are differentiated by various indices
such as temperature, humidity, radiation, heat, etc. Further discussion of this topic is beyond the scope of this introductory
chapter but the point is that substances can have inherent hazards of varying degrees and this influences the risks (Equation
1.1). An adequate hazards assessment is vital and should be
done by knowledgeable individuals.
1.6
EXPOSURE CHARACTERISTICS
Exposure involves a hazard source, a pathway and a target
or receiver. Industrial ventilation controls are used to separate
the target or receiver from the airborne hazard source by reducing the likelihood of exposure. Exposure to a hazard is related
to the amount (How much?); the frequency (How often?), and
the duration (How long?). With regards to health hazards,
“Dose” is used to describe the amount of exposure. Dose is
how much gets in or on the body. An exposure assessment is an
attempt to estimate the dose to workers or other individuals
using various methods and approaches (Section 1.7).
Lack of adequate exposure controls obviously dictates the
likelihood of exposure. However, when conducting risk
1-8
Industrial Ventilation
FIGURE 1-1. Evaporation of volatile liquid
FIGURE 1-3. Dust generated from falling materials
assessments and assessing the need or adequacy of controls,
there are certain circumstances and/or characteristics of materials that can also provide an indication of the likelihood of
exposure to airborne hazards (Table 1-7). Some have already
been alluded to in previous sections but are important enough
to be mentioned again. Sometimes predisposing conditions are
also required for a risk to be realized (e.g., flammable or combustible materials require a certain concentration in air and
ignition source for fire or explosion). Often, these conditions,
circumstances, and/or characteristics can be manipulated using
the hierarchy of controls to reduce the hazards and/or reduce
the likelihood of exposure and minimize the risks (Section
1.8).
likelihood of exposure (Section 1.7). Amount is not only related to the quantity of material available to become airborne but
the concentration of hazardous material in the air. Conditions
of dispersion, physical form, volatility, aerodynamic diameter
and moisture content are some of the major factors that influence the likelihood of a material becoming airborne. Substituting large amounts of hazardous materials with smaller
amounts in a process is a simple method to reduce the likelihood of exposure and the risks.
1.6.2 Conditions of Generation and Dispersion. As mentioned previously, energy is required to generate and disperse
1.6.1 Amount, Duration and Frequency. The greater the
amount, duration and/or frequency of exposure, the greater the
FIGURE 1-2. Dust expulsion by mechanical compression
FIGURE 1-4. Displaced air containing fine particulate
Risk Assessment
1-9
indicator of volatility. The lower the boiling point, the greater
the volatilization and the greater the likelihood of exposure.
Also, the more surface area a volatile substance has exposed to
air, the greater the amount of volatilization (Section 1.6.1).
1.6.5 Aerodynamic Diameter.(1.5) Aerosols require energy
to become airborne and settle out of the air by gravity, based
on their aerodynamic diameter. The aerodynamic diameter is
the diameter of a hypothetical sphere of density 1 g/cm3 having
the same terminal settling velocity in calm air as the particle in
question, regardless of its geometric size, shape and true density. Aerodynamic diameter will determine if, and for how
long, an aerosol remains airborne. With regards to health
effects on individuals, it also determines the likelihood of the
substance being inhaled and the site of deposition in the respiratory tract.
FIGURE 1-5. Dust created by abrasive blasting
airborne hazards (Section 1.5.2). Dispersion can happen as the
airborne hazard is generated or, in the case of solid aerosols,
can also occur when the solid is re-aerosolized after settling.
The greater the energy input into a process, the greater the
potential for airborne hazards to be generated and dispersed.
Increasing heat, motion, energy of mechanical breakdown, fall
height of material, pressure changes causing movement of fluids (gases and liquids) and rate of chemical reactions, all
increase the likelihood of exposure (Table 1-3).
1.6.3 Physical Form. In general, smaller more fluid substances are more likely to become airborne and stay airborne,
where there is a greater likelihood of exposure. For example,
airborne hazards in the forms of gases, vapors, smoke, fogs,
fume and fine powder are more likely to result in exposure
than larger aerosols such as coarse powders or large droplets.
Volatility is a descriptive characteristic of fluids and some
solids and can be used to predict the likelihood of a material
becoming airborne. Aerodynamic diameter is a descriptive
characteristic of aerosols, which can be used to predict the
likelihood of a material becoming airborne and staying airborne. Substituting hazard forms that are less likely to become
airborne is an effective risk control technique.
1.6.4 Volatility. Vapors and gases move from their source
and through the air by diffusion from high concentration to
low concentration based on their vapor pressure. The higher
the vapor pressure, the faster the diffusion and the greater the
likelihood of exposure. Vapor pressure is temperature dependent and increases with increasing temperature (i.e., energy
into process). Substances having higher vapor pressures are
therefore more difficult to contain and control. When vapor
pressure data are not available, boiling point can be used as an
The larger the aerosol, the faster it settles out of the air.
Particles with an aerodynamic diameter larger than 100 µm
can become airborne, depending on conditions, but barely
remain in the air. Particles larger than 50 µm do not remain airborne very long; their settling velocity is greater than 7 cm/sec.
Small aerosols (less than 1 µm) such as smoke, fumes and
nano-size particles remain airborne for much longer periods.
They have a settling velocity of about 0.03 mm/s and tend to
behave more like gases and vapors and move with air currents.
Aerosols that are small enough to stay airborne may be
inhaled. Aerosols with an aerodynamic diameter greater than
30 µm are deposited predominantly in the airways of the head.
Larger aerosols, which do not deposit in the head, will be
deposited in the airways of the lungs. Particles smaller than 10
µm may penetrate the alveolar regions of the lungs, where
inhaled vapors and gases can be absorbed. Fibers, because of
their geometry, tend to have a smaller aerodynamic diameter
than their length would suggest. Because of this, long fibers
may also reach the alveolar regions of the lungs.
1.6.6 Moisture Content. Wet or damp materials are less
likely to release solid aerosols (e.g., dust or fibers) and result
in exposure. In materials that are not initially wet or damp,
moisture can be added to prevent or minimize aerosols.
Moisture can also be used to suppress or capture dust or fibers
that are already airborne. Water sprays, with or without chemical additives such as surfactants, create water drops that collide with airborne particles. The moisture increases the particle’s aerodynamic diameter causing it to settle out of the air
faster. Addition of moisture is a common practice in construction and demolition work for controlling exposure to asbestos,
lead and crystalline silica.
1.7
HEALTH HAZARD EXPOSURE ASSESSMENT(1.3)
Inhalation exposure to airborne health hazards is a main concern in relation to industrial ventilation. An exposure assessment is an attempt to quantify or estimate the amount, duration
and frequency of such exposures. These exposures can be
quantified or estimated using various methods and approaches;
1-10
Industrial Ventilation
TABLE 1-6. Airborne Hazard Categories and Types
and results can be compared to applicable exposure limits to
judge the acceptability of exposures. Occupational exposure
assessments should be conducted by qualified industrial
hygienists or other competent professionals who are trained in
assessing occupational exposures. Prior knowledge of the hazards and the severity of their potential harm (Section 1.5) are
prerequisites in an exposure assessment, as the two are integral
parts of the risk assessment process.
1.7.1 Observations and Monitoring. Observations may be
used to identify conditions of airborne hazard generation and
dispersion (Section 1.6.2) and determine where monitoring
might be appropriate. Monitoring is the process of evaluating
and documenting exposures to existing airborne hazards.
Monitoring can be qualitative, semi-quantitative or quantitative.
Exposures can sometimes be sensed visually or by odor and
irritation; approximated by screening measurements or “grab
samples,” or if necessary, more closely defined with real-time
and/or time-integrated air monitoring. Visually-apparent airborne hazards (e.g., dust clouds) are usually a sign that exposures are excessive. However, the lack of visual indicators
does not mean exposures are absent or insignificant. Many airborne hazards are invisible. For hazards having odor and/or
producing irritation, these sensations can be compared to published data to estimate general exposure levels. Grab sampling
can be used to collect a snapshot of exposures at a particular
time and place. These methods can be used to assess worstcase exposure conditions and employed as screening tools for
judging acceptable and unacceptable exposures.
Direct-reading instruments can be used to quantify airborne
Risk Assessment
1-11
TABLE 1-7. Likelihood of Exposure Indicators for Airborne Hazards
hazard concentrations in real-time to assess peak exposures,
pinpoint locations of generation and establish relationships
between activities and exposure levels. When indicated,
direct-reading instrument use may be followed up with more
specific exposure measurements, which are more indicative of
ambient work area concentrations and/or breathing zone exposures. Personal (breathing zone) exposure monitoring is the
best way to assess a worker’s exposure to airborne health hazards. Area (ambient) samples represent the airborne concentration in a specific location. Time-integrated air monitoring techniques can be used to collect airborne contaminants on media
for later analysis at an accredited laboratory. These methods
can be used to assess average exposures over a particular time
period. The time-weighted average (TWA) exposures can be
standardized and directly compared to established, occupational exposure limits or OELs (Section 1.7.4).
Direct-reading instruments can also be combined with
video use to more clearly visualize conditions and sources of
exposures. Video exposure monitoring (VEM) is a technique
where worker exposures are monitored with direct-reading
instruments while workplace activities are simultaneously
recorded on videotape. The method is particularly useful in
identifying where the majority of the exposure occurs and
where controls could be most effective.
1.7.2 Surrogate Data. Objective data based on past monitoring, industry-wide surveys and monitoring data from other
comparable agents and operations can also be used to predict
worker exposures. Unless the particular process, task, activity,
material and/or exposure situation in question is unique, similar or comparable circumstances have likely been assessed
before. That information may be available, relevant and useful.
Primary scientific literature, governmental and professional
organizations, manufacturers, unions, trade organizations, etc.,
can be useful exposure assessment data sources. Data considered should be from a credible source and should closely represent the particular exposure conditions being considered or
should be interpreted with caution.
1.7.3 Modeling. Modeling is a tool for predicting exposures
by mathematically exploring hypothetical exposure situations
based on input data and conservative assumptions regarding
hazardous agents and/or workplace conditions. Modeling is
often used as a screening tool to estimate current exposures,
predict potential future exposures or re-create historical exposures.
Exposure models range from relatively simple, straightforward tools to provide a quick, rough approximation of exposure (see examples in Chapter 10), to more detailed, sophisticated methodologies that can refine the exposure estimate statistically (Chapter 12). The former are useful as screening
tools for judging acceptable exposures and prioritizing efforts
for gathering further exposure information. The latter are often
used to more accurately interpret limited monitoring data.
1.7.4 Occupational Exposure Limits. An Occupational
Exposure Limit (OEL) is an average exposure, for a length of
time, which defines an upper limit on the acceptable workplace
1-12
Industrial Ventilation
exposure for a specific airborne hazard or class of hazards.
OELs are based on toxicological and epidemiologic health
effect data. Ideally an OEL indicates an exposure level below
which it is believed most workers will not experience adverse
health effects. The averaging exposure times most often used
for OELs are 15-minutes (short-term exposure limits or
STELs), 8-hours (8-hr TWAs) and 10-hours (10-hr TWAs).
However, some OELs are instantaneous (ceiling) limits.
alone. Exposure estimates based solely on surrogate data
should err on the side of caution and be considered conservatively. In the end, unacceptable exposures should be controlled
(Section 1.8) and further information should be gathered on
uncertain exposures to make a better judgement on their
acceptability.
Formal OELs exist for a relatively small number of commonly-used industrial chemicals. They have been established
primarily by governmental agencies and/or non-governmental
authoritative groups (Table 1-8). In some cases, chemical manufacturers have established internal OELs when regulatory
and authoritative OELs are absent. Many countries have OELs
that are enforced by legislation to protect occupational health.
Looking at Equation 1.1 (Section 1.2), it is clear that risks
can be avoided by eliminating the hazard(s) or controlled by
reducing the hazard(s) and/or reducing the probability of exposure. This is done by considering the hierarchy of controls
(Figure 1-6). The hierarchy of controls is a sequential ranking
and order for considering and implementing various controls,
from most effective to least effective, to eliminate or minimize
risks. Control methods at the top of the hierarchy are potentially more effective and protective than those at the bottom.
Allocating time and resources to control risks up front by considering the hierarchy of controls gives a much better return on
investment and a more sustainable solution.
Formal OELs do not exist for the vast majority of hazardous
chemicals used in industry. In these cases, working OELs, or
informal limits, are established using what little scientific data
are available. Some working OELs are established by analogy
with similar agents having formal OELs. Due to the lack of
available data and uncertainty, working OELs are typically
stated as exposure ranges, or bands, as opposed to values.
1.7.5 Assessing Exposures. The industrial hygienist or
other professional must compare the exposure estimate and its
uncertainty to the chosen OEL and its uncertainty. They must
use available qualitative and quantitative data, professional
judgement and statistical tools to judge exposures as acceptable, unacceptable or uncertain.
Certain industrial processes are known to produce unacceptable exposures and require controls (Table 1-5). Visual observations and/or odor and irritation are sometimes sufficient
information for an experienced industrial hygienist to make an
exposure assessment judgement on clearly acceptable and
unacceptable exposures. For questionable exposures, quality
monitoring data should most influence the exposure assessment judgement, especially when combined with statistical
modeling methods. Monitoring and objective data should be
given more weight in judging exposures than modeling data
1.8
HIERARCHY OF CONTROLS
1.8.1 Elimination or Substitution. Initial efforts to address
risks should focus on risk avoidance through hazard elimination by physically removing the hazard. If not possible, hazard
substitution should be considered by replacing the hazard with
a less hazardous material or process. An example is the
replacement of solvent-based paints with water-based paints or
dip coating materials rather than spray coating to reduce the
possibility of exposure. This should be followed by the application of engineering controls that are supplemented by
administrative controls. Personal protective equipment should
only be considered when other controls are not technically,
operationally or financially feasible or during the installation
of other control measures. If risks cannot be avoided by elimination, often a combination of methods in the hierarchy may
be necessary to control the risks.
1.8.2 Engineering Controls. Engineering controls are used
to isolate people from hazards and deter worker error by the
installation of equipment, or other physical facilities.
TABLE 1-8. Commonly-used Formal Occupational Exposure Limits in the United States
Risk Assessment
1-13
FIGURE 1-6. Hierarchy of Exposure Control Measures(1.4)
Examples include containment, enclosures, barriers, interlocks
and ventilation – the subject of this Manual. Ventilation is used
to control airborne hazards by exhausting or supplying air to
either remove hazardous atmospheres at their source or dilute
them to a safe level. The two types of ventilation are typically
termed local exhaust and general or dilution ventilation. Local
exhaust attempts to enclose the material, equipment or process
as much as possible and to withdraw contaminated air at a rate
sufficient to assure that the direction of air movement at all
openings is always into the enclosure. General or dilution ventilation attempts to control hazardous atmospheres by diluting
the atmosphere to a safe level by either exhausting or supplying air to the general area. Local exhaust is normally the ventilation method of preference, as it is more effective.
Engineering controls tend to be more effective than administrative controls (Section 1.8.3) and personal protective
equipment (Section 1.8.4) because they are less dependent on
the worker. Workers, unfortunately, are subjected to all of the
frailties that befall humans (e.g., forgetfulness, preoccupation,
insufficient knowledge).
1.8.3 Administrative Controls. Effective engineering controls require the application of administrative controls as either
supplemental hazard controls or to ensure that the engineering
controls are developed, maintained, and properly functioning.
Administrative hazard controls consist of managerial efforts to
reduce risks. These efforts can include: planning, information
and training (e.g., hazard communication), written policies and
procedures, safe work practices, and environmental and medical surveillance (e.g., workplace inspections, equipment preventive maintenance, and exposure monitoring). Because they
primarily address the human element of hazard controls, they
are of vital importance and are always used to control hazards.
1.8.4 Personal Protective Equipment. Personal protective equipment (PPE) includes a wide variety of items worn
by an individual worker to isolate the person from hazards.
PPE includes articles to protect the eyes, skin, and the respiratory tract (e.g., goggles, face shields, coats, gloves, aprons,
respirators). In some situations, PPE may be the only reasonable hazard control option, but for many reasons it is the least
desirable means of controlling hazards and should be the last
control choice considered. PPE does not eliminate hazards
but merely reduces the probability of exposure. The effectiveness of PPE is highly dependent on the user. PPE is oftentimes cumbersome and uncomfortable to wear. Each type of
PPE has specific applications, advantages, limitations, and
potential problems associated with misuse and those using
PPE must be fully knowledgeable of these considerations.
PPE must match the hazards and the conditions of use and be
properly maintained in order to be effective. Misuse may
directly or indirectly contribute to the risks being addressed or
create new risks. The material of construction must be compatible with the hazards and must maximize protection, dexterity, and comfort.
1.9
SUMMARY
The core of the risk assessment process involves the following steps: 1) anticipate, identify and analyze all the hazards,
including airborne hazards; 2) assess the exposures; 3) estimate the risks, and 4) for any unacceptable risks, determine the
appropriate controls (Figure 1-7). Industrial ventilation is used
to decrease the likelihood of exposure to airborne hazards. If
ventilation controls are determined to be the appropriate haz-
1-14
Industrial Ventilation
FIGURE 1-7. Summary of risk assessment process with emphasis on airborne hazards
Risk Assessment
ard control and implementation is warranted based on the risk
assessment, the information gathered in the risk assessment
will be valuable not only to the design process , but in installation, operations and maintenance, as well. This Manual, and
the companion Operation and Maintenance Manual, provide
the details in order to accomplish these tasks.
REFERENCES
1.1
1.2
U.S. Department of Defense: Standard Practice for
System Safety. MIL-STD-882E (2012).
American National Standards Institute/American
Society of Safety Engineers: Prevention Through
Design Guidelines for Addressing Occupational
Hazards and Risks in Design and Redesign Processes.
American National Standard Z590.3. ANSI. Des
Plaines, IL (2011).
1-15
1.3
American Industrial Hygiene Association: A Strategy
for Assessing and Managing Occupational Exposures.
AIHA, Falls Church, VA (2015).
1.4
National Institute for Occupational Safety and Health
(NIOSH) website http://www.cdc.gov/niosh/topics/
hierarchy/default.html (2017).
1.5
World Health Organization: Hazard Prevention and
Control in the Work Environment: Airborne Dust.
WHO, Geneva (1999).
1.6
United Nations Economic Commission for Europe
(UNECE): Globally Harmonized System of Classification and Labeling of Chemicals (GHS), Sixth
revised edition, UN, New York and Geneva (2015).
1.7
American Industrial Hygiene Association: The
Occupational Environment – Its Evaluation and
Control. AIHA, Fairfax, VA (1998).
Chapter 2
PRELIMINARY DESIGN AND COST ESTIMATION
NOTE: Equations with notation followed by (IP) are designated for inch-pound system only; equations followed by (SI) are designated for
metric use only. If equation bears neither, then it applies to both systems.
2.1
2.2
INTRODUCTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-2
PROJECT GOALS AND SUCCESS CRITERIA . . . . .2-3
2.2.1 Small Projects or Small Organizations and
Success Criteria . . . . . . . . . . . . . . . . . . . . . . . . .2-3
2.2.2 Larger Projects and the Keys to Success . . . . . .2-3
2.3 LARGE PROJECT TEAM ORGANIZATION . . . . . . .2-4
2.4 TEAM RESPONSIBILITY MATRIX (TRM) . . . . . . . .2-4
2.5 PROJECT TEAM SAFETY . . . . . . . . . . . . . . . . . . . . . .2-8
2.5.1 Process and Equipment Safety Studies . . . . . . .2-8
2.6 DOCUMENT CONTROL . . . . . . . . . . . . . . . . . . . . . . .2-8
2.7 PROJECT TEAM ORGANIZATION, SELECTION
AND SKILLS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-8
2.8 INTERNAL RESPONSIBILITY FOR FINAL
APPROVAL OF BUDGET, TECHNICAL MERIT
AND REGULATORY ISSUES . . . . . . . . . . . . . . . . . . . .2-9
2.9 COMMUNICATION OF PLANT (AND PROJECT)
REQUIREMENTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-9
2.9.1 Project Feasibility and Conceptual Design . . . .2-9
2.9.2 Design Definition – Defining and
Communicating the Scope . . . . . . . . . . . . . . . . .2-9
2.9.3 Detailed Design . . . . . . . . . . . . . . . . . . . . . . . .2-11
2.10 DESIGN/BUILD, IN-HOUSE DESIGN OR
OUTSIDE CONSULTANT . . . . . . . . . . . . . . . . . . . . . .2-11
2.11 DESIGN-CONSTRUCT METHOD (SEPARATE
RESPONSIBILITIES FOR ENGINEERING AND
INSTALLATION) . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-14
2.11.1 Selection of Engineering Firm . . . . . . . . . . . . .2-14
2.12 DESIGN/BUILD (TURNKEY) METHOD –
SINGLE SOURCE OF RESPONSIBILITY . . . . . . . .2-16
2.13 PROJECT TEAM AND SYSTEM EVALUATION . . .2-16
2.14 PROJECT RISK AND NON-PERFORMANCE . . . . .2-17
2.14.1 Communication of Risk . . . . . . . . . . . . . . . . . .2-17
2.14.2 Communicating Proof of Performance . . . . . .2-18
2.15 USING PLANT PERSONNEL AS PROJECT
RESOURCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-18
2.16 INTERFACE BETWEEN THE PLANT AND
PROJECT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-18
2.17 IMPACT OF NEW SYSTEMS ON PLANT
OPERATION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-19
2.18 CAPITAL COSTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-19
2.18.1 Order of Magnitude Pricing Methods . . . . . . .2-19
2.18.2 Equipment Factor Pricing Method . . . . . . . . . .2-20
2.18.3 Price Book Estimating Method . . . . . . . . . . . .2-20
2.19 OPERATING COST METHODS . . . . . . . . . . . . . . . . .2-21
2.19.1 Annual Operating Cost Components . . . . . . . .2-21
2.19.2 Estimation of Total Annual Cost . . . . . . . . . . .2-22
2.19.3 Factored Methods for Determining Annual
Costs . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-22
2.19.4 Annualizing the Capital Cost of a Project . . . .2-22
2.20 COST COMPARISON METHODS . . . . . . . . . . . . . . .2-23
2.20.1 Simple Payback Period Method . . . . . . . . . . . .2-24
2.20.2 Abbreviated Life Cycle Cost Method . . . . . . .2-24
2.20.3 Life Cycle Payback Method . . . . . . . . . . . . . .2-24
2.20.4 Life Cycle Costing Considering Taxes . . . . . .2-25
REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2-26
____________________________________________________________
Figure 2-1
Figure 2-2
Figure 2-3
Figure 2-4
Sample Team Responsibility Matrix . . . . . . . . . . .2-5
Figure 2-5 Equipment Factored Estimates . . . . . . . . . . . . . .2-20
Sample Project Closure Document (PCD) . . . . . .2-7
Figure 2-6 Suggested Factor Unit Cost: Thermal
Sample Design Basis Form . . . . . . . . . . . . . . . . .2-10
Fluid-Bed Catalyst Unit . . . . . . . . . . . . . . . . . . . .2-22
Sample Design Basis . . . . . . . . . . . . . . . . . . . . . .2-12
____________________________________________________________
Table 2-1
Project Design and Construction Methods –
Impact on Owner . . . . . . . . . . . . . . . . . . . . . . . . .2-15
Operating and Maintenance Requirements . . . . .2-21
Table 2-2
Table 2-3
Table 2-4
Life Cycle Payback Analysis . . . . . . . . . . . . . . . .2-25
Example Problem 2-10, Life Cycle Payback
Analysis with Tax Implications . . . . . . . . . . . . . .2-26
NOTE: The section covering cost estimating uses mostly IP units because of the variety of cost values that could be used worldwide. Where
appropriate, metric values are inserted but example problem solutions are in IP units only.
2-2
2.1
Industrial Ventilation
INTRODUCTION
When contemplating preliminary design and cost of a ventilation system first consider the type of system required.
There are two generic types of ventilation systems normally
used in industrial plants – air supply and exhaust systems.
Exhaust systems are used to remove contaminants generated
by manufacturing operations and assist in maintaining a
healthy environment. Supply air systems replace the exhaust
air and/or provide heating, ventilating and air conditioning.
Generally, when an exhaust ventilation system is installed, it
must be accompanied by a supplied air system to replace the
exhausted air. If more air is exhausted from the workspace
than the quantity of air supplied, air will enter the facility in an
uncontrolled manner through cracks, walls, windows, doorways and possibly through flue pipes.
Many times exhaust and supply ventilation systems are
installed to reduce the risk of employee exposure to a hazardous air contaminant by controlling emissions of toxic materials. The complexity and cost of the ventilation system will
depend on the size of the system and the amount of contaminant control required. For a more toxic air contaminant
(greater health hazard potential), a higher level of control is
needed to maintain the contaminant concentration below the
desired or regulated level. A higher level of control can mean
larger volumes of exhaust and supply air, larger air handling
systems, more elaborate hoods, etc., thus, more intricate
design and higher cost.
The toxicity of hazardous air contaminants is characterized
by what is considered a safe amount of material an employee
can be exposed to for a length of time. Regulatory and nonregulatory agencies publish occupational exposure limits that
rate the relative toxicity of chemicals (see Chapter 1, Section
1.7.4 and Table 1-8).
The preliminary design of a well-designed exhaust system
will consider: if and which OELs are targeted; the volume of
air needed to control the hazards to or below the selected OEL;
the contaminant collection hoods necessary to provide capture; the duct sizes necessary to maintain minimum transport
velocities through the system; the air cleaning equipment necessary to clean the captured air; fan and motor requirements;
the location of the air cleaning equipment, fan and motor;
required safety devices to minimize dust deflagration hazards;
the maintenance needed to assure continuing performance;
and safety equipment (guarding, fall protection, PPE) needed
by the maintenance staff to safely maintain the system, etc.
determined in the planning stages of a project, cost estimating
methods and guidelines can be used to predict these costs with
some degree of accuracy.
The same project can have a number of different cost estimating requirements depending on what type of information is
needed. Management may want the total capital outlay for a
project (before the project starts and during the installation) and
may need to determine the cost of downtime for production
during construction. Maintenance will need to know the cost in
labor hours for start-up, training and operation. Plant planning
personnel will need information on annual operating expenses,
utilities and manpower needs. The engineering manager will
need to allocate labor-hours and cash for design services, construction supervision and technical services. Environmental
regulators may ask for a project cost estimate as part of a Best
Available Control Technology (BACT) analysis or other permit
application requirements.
Some costs may be simple to estimate. A catalogue item
may have a fixed and predictable price. The cost of custom
built equipment and fabricated duct components can vary significantly from standard catalogue offerings. Special equipment such as fire or explosion protection can add significant
cost to a system and should not be overlooked in developing
the cost estimate. The operating and maintenance costs for the
system can also be difficult to estimate with accuracy before
the details of the system are known. The price of equipment
can vary significantly depending on the level of competition
between vendors and market conditions for raw materials.
Installation costs will vary depending on site conditions and
the availability of local construction resources. Effective construction planning and management are also critical for managing installation costs and staying on budget.
The accuracy of the cost estimate will increase as a project
progresses from the initial concept planning phase through to
construction as the specifics of the system become more
defined. During the initial phase of a project, only the emission
sources will be known. Since the design details of the industrial
ventilation system are not known, ‘rule of thumb’ system cost
estimates are commonly applied. The United States
Environmental Protection Agency (USEPA) and other sources
have developed methods for making rough estimates on control costs based on the process type.(2.1) The level of accuracy
can be very low (± 50%), but it can be enough to begin to determine basic cost information and develop a preliminary budget.
The preliminary design of a well-designed supply system will
supply the volume of air needed for replacement and consist of
an air inlet section, filters, heating and/or air cooling equipment,
fan/motor, and ducts and registers/grill for air distribution.
During the preliminary design phase, the design team will
conduct a study to determine the level of control required, control options, potential emission sources, control equipment
locations, and estimates of air volumes. This information can
be used to refine the cost estimate to within a factor of ± 30%.
An industrial ventilation system will have many costs associated with its construction and operation. Accurate cost estimating is important to the success of a ventilation project.
Although the exact price tag for most installations is not easily
Once detailed engineering has been completed, a detailed
cost analysis can be made. Depending on the method and
accuracy of the estimator, this can be within ± 5% of the project cost.
Preliminary Design and Cost Estimation
In this chapter, the basics of designing a system and developing a project budget, determining economic advantages of
multiple control options, estimating annual operating and
maintenance costs, and reporting of costs to environmental
agencies will be covered. Detailed cost estimating is a highly
specialized task that is beyond the scope of this Manual.
2.2
PROJECT GOALS AND SUCCESS CRITERIA
Because the design and installation of a local exhaust ventilation system involves approval by many outside agencies and
potential interface with varied plant processes and departments,
management must select a team that is responsive to identifying
the end user(s) and their needs. In most cases, errors can occur
between the company and regulator, the company and contractor or among project team members themselves. Communication and organization are necessary for a successful installation.
Team members must have or develop these skills to bring success to the project (see Chapter 4, Figure 4-1).
2.2.1 Small Projects or Small Organizations and Success
Criteria. Simple projects and smaller operations may not need
a team or special organization to complete the installation of a
ventilation system. However, even the smallest system has
requirements to meet safety and environmental regulations.
Because of these regulatory aspects, all projects should have
minimum organization and documentation.
2-3
Commercial requirements such as funding, cost controls and
budget management would be required.
Even the smallest systems require a review on the impact to
existing plant resources such as electrical power, floor space
and production and maintenance staff. This can include some
preliminary engineering from vendors or engineering staff to
provide a concept of the design. Production disruptions and
adjustments should also be identified and planned. In general,
every project could be organized in the following phases:
1) Feasibility and concept design – The idea the process is
defined, studied and verified for the new project. Sizes
for equipment are estimated and preliminary estimates
are made for feasibility. The issues of Proof of
Performance and Commissioning could be included at
this stage.
2) Definition and funding – The design is refined so that
the scope can be written for instructions to designers or
for design/build firms. Detailed (± 20%) estimates are
made so that funding can be acquired, and work
required to start permit process is determined.
3) Detailed design – This should be done either by inhouse, independent or design/build engineering staffs
with enough detail to evaluate the process impacts and
cost issues.
All ventilation projects usually begin with problem identification and should conclude with proof of performance (see
Chapter 2 of Industrial Ventilation: A Manual of Recommended Practice for Operation and Maintenance [the O&M
Manual]). Ventilation problems can be in response to the following issues:
4) Construction – This can begin during the detailed
design phase after design approval and securing of
required permits.
1) Addition of new process that requires ventilation controls;
2.2.2 Larger Projects and the Keys to Success. As sys-
2) Change to existing process that requires additions or
revisions to existing systems;
3) Measured or perceived safety and health issues that can
be improved with ventilation;
4) Response to plant labor committees to improve ventilation for worker comfort or safety;
5) Needed improvements to a poor design that rendered
the system to be ineffective or to cause a waste of energy;
6) Failure of the present system to meet required emission
levels or control an air contaminant to present OELs;
7) Become compliant with current OELs.
In response to the needs for ventilation system installation or
improvement, management must mobilize the necessary plant
resources. In a small plant this may be just the plant manager or
engineer working with outside contractors or engineers.
Instructions for installation may be given verbally or with a few
sketches and followed with a formal or informal proposal.
5) Startup and Commissioning – This is final phase where
ownership of the project is transferred from the construction organization to the final owner.
tems become larger and more complicated and/or have implications for meeting regulations, the need for more formal
organization and document controls becomes mandatory. This
includes the documentation of the basis for design, the conceptual and detailed design drawings and verification of effectiveness of the system. In addition, maintenance and service
records for the completed installation need to be kept. This is
required for a proper transfer of ownership from the project
team to the system operators (see Chapter 2 of the O&M
Manual).
The same goals and phases also govern smaller projects, but
the organization may be less formal for small projects.
However, communications, especially among team members,
should be documented.
Organization starts with the identification of the person or
persons responsible to receive and operate the system.
Maintenance and production may be assigned responsibility
for the upkeep of the supply and exhaust systems and replacement parts, and must be kept informed throughout the design
process.
2-4
Industrial Ventilation
The first step is a simple document to define the expectations (success criteria) for the operation of the system. The
expectations become the directive to determine all further
effort. In its simplest form, this document notes existing problems or shortcomings to be resolved, and can be in long hand
or outline form. These concerns could include exceeding
OSHA limits, poor performance of existing control technology, etc. Other advisory groups such as The American
Conference of Governmental Industrial Hygienists
(ACGIH®), The American Industrial Hygiene Association
(AIHA) and the National Fire Protection Association (NFPA)
can provide supplemental data for occupational health and
safety exposure limits. In some cases, it may be good to refer
to previous projects and what had been considered successful
project completions. Evaluating unsuccessful projects may
also provide insight into deficiencies to avoid. This may be
done by looking at a design basis (see Section 2.9.2) to reference the closure requirements. Looking at the end of project
requirements can better define what needs to be accomplished
at the beginning of a new installation.
The document would then identify measurable goals (dust
exposure, emission levels, bringing a process on line by a certain date, etc.) so that plant management, design engineers and
contractors stay focused on the system requirements. It would
also identify the benefits (cost or energy savings, avoidance of
fines, etc.) so that the clear intent of the installation is maintained. Any system, no matter how small, that includes responsibility to regulatory agencies and has impact on worker health
and safety should include this important first step.
Evaluation should also include an assessment of potential
risks (see Chapter 1). This includes evaluation of potential
risks for worker exposure to the dusts, mists, fumes, vapor or
heat from the process. These risks primarily include inhalation
but may include other exposures such as skin absorption, skin
irritation, or allergic response. This document also may
include input from manufacturers of new or existing equipment or processes, to see if there are alternative methods to
reduce exposure. Before the scope is defined, it must be determined that all practical means have been investigated to
remove or reduce the contaminants at their source before
adding controls.
Any restrictions on the industrial ventilation (IV) system or
controlled process should be listed. This may include access to
equipment that is hindered by hoods or ergonomic considerations (e.g., as workers need to reach into enclosures or over
other restrictions from the system). Maintenance and construction worker access and safety must also be identified. Several
of these publications are listed in the references section.(2.2–2.7)
Use the most current information when preparing an estimate.
Some of the listed sources provide annual updates of the costs
provided, and the pricing books have adjustment factors to
modify labor cost according to the location of the project.
During this early stage the need for environmental permits
must be addressed. In many states this process can take more
than a year and potentially delay the construction or start-up.
If sufficient resources are not available within the company,
then a permit specialist (consultant or law firm) should be contacted early in the project schedule.
The Clean Air Act Amendments of 1990 have changed
many of the requirements, especially with issues such as
Maximum Achievable Control Technology (MACT)
Standards (Title III), permits (Title V), non-attainment areas,
permits to install and permits to operate, etc. In many cases,
preliminary estimates are done with regard to engineering data
(emission factors, air volume, stack heights and locations, etc.)
that must be accomplished even before organizing a project
team. Similarly there may also be a need to determine if studies are needed for Prevention of Significant Deterioration
(PSD) permits and if best available control technology or other
provisions are needed. These early reviews may actually provide opportunity to save on the installation by considering
alternate processes or materials to eliminate or reduce the need
for pollution controls.
2.3
LARGE PROJECT TEAM ORGANIZATION
After project goals have been determined and a person is
designated to receive the finished project, the task of organizing within the plant begins. Again, the size of the project may
determine the experience needed to proceed.
At a minimum, representatives from Process, Purchasing,
Occupational Safety & Health (OSH), Maintenance and Plant
Engineering are required (Figure 2-1). In smaller operations,
one or two persons may hold all these positions. In addition,
there probably are requirements for approval by regulatory
agencies (requiring stack testing for emissions or industrial
hygiene testing for OSHA issues). These health and safety
reviews by plant professionals or consultants should be included for any system that has an impact on the plant environment.
While workstation operators do not have to be part of the
larger design team, they must be consulted regularly during the
design process and may have suggestions that may make the
project work smoother. For example, mockups of hood and
enclosure designs can define operability problems that can be
addressed during the design phase.
2.4
TEAM RESPONSIBILITY MATRIX (TRM)
At this point an outline of team member responsibilities
should be developed (Figure 2-1). This outline is called a Team
Responsibility Matrix or TRM, and can also include the
requirements of outside resources such as consultants, e.g.,
industrial ventilation design specialists or special service companies (industrial hygiene firms, etc.). At the same time, the
Project Closure Document (Figure 2-2) should be completed
to determine the persons responsible for final acceptance of
the project. Some preliminary work can also be accomplished
for the commissioning process, such as a list of proofs of per-
Preliminary Design and Cost Estimation
FIGURE 2-1. Sample Team Responsibility Matrix
2-5
2-6
Industrial Ventilation
FIGURE 2-1 (Cont.). Sample Team Responsibility Matrix
Preliminary Design and Cost Estimation
FIGURE 2-2. Sample Project Closure Document (PCD)
2-7
2-8
Industrial Ventilation
formance. These could include items such as required filter
bag life, emission levels from the collector, TLV® near operator station(s), etc. Typical plant personnel to be included are
shown on the form, but may be expanded based on particular
project needs. See Chapter 2 of the O&M Manual for a complete discussion on commissioning and system evaluation.
The purpose of the TRM is to ensure that the proper
resources are used to determine the plant and project needs
before the design begins. The boxes on the form would contain
the names of the individuals responsible for the input to the
Design Basis (instructions to the design team) and the project.
The individuals would initial opposite their name to indicate
that the information has been given to the Project Manager for
issue. These same individuals would initial in the remaining
boxes after issuance of the Design Basis and the construction
package (instructions to the contractors and/or bidders). This
minimizes delays and scope changes as the project proceeds.
It also avoids late input from outside sources that could
impede the project timing and success.
2.5
PROJECT TEAM SAFETY
Prime considerations when beginning these projects are
health and safety. This includes the safety of the audit process
since readings of pollutant and energy outputs may be required,
and extends both to plant workers and outside testing and engineering firms. The data required for the design of air pollution
control or industrial ventilation systems may not be normal
measurements taken in the process. Special plant precautions
may be required to manage the safe gathering of information.
For instance, many air pollution control systems require
scaffolding to perform source emissions acceptance tests.
Initial testing, adjusting and balancing technicians may need
cherry picker trucks to access sub-main ducts located over
some processes. The contract must be written to inform them
of the safety needs, such as appropriate scaffold and scaffold
platform design, and respiratory and fall protection.
2.5.1 Process and Equipment Safety Studies. Similarly,
the attachment of air pollution control devices to existing
processes may have impacts on the processes themselves.
Process safety reviews may be necessary to evaluate the
impact of the system additions. For instance, the purchasing
department may need to locate sources and storage facilities
for treatment chemicals, or filter media. Wet collectors may
require additional permitting to discharge into the industrial
waste treatment plant or sewer system. Personnel safety or fire
and explosion studies may be required based on the nature of
the project. Debris collected in material air separators may
require explosion severity (Kst) testing to determine if it is a
combustible dust.
2.6
DOCUMENT CONTROL
Smaller projects may have little in the way of drawings,
specifications or design calculations. Because ventilation proj-
ects may have regulatory or safety implications, there should
be some record of the system design and maintenance requirements.
The control of project documents begins immediately. For
larger projects this includes minutes of planning meetings,
meetings with contractors and consultants and the exchange of
information such as scope of work and bid proposals. The document control may be as simple as a correspondence file kept
by the plant engineer or project manager.
Document control can also serve to keep the project
focused. Often a project is expanded as other plant needs are
addressed. This is not always negative as pollution control
projects often can be the opportunity to improve plant efficiencies and reduce operating costs. Document control can be used
to manage the input for scope definition (scope creep) and
resultant increased project costs.
Document control can also be invaluable for the avoidance
and settlement of project disputes. Many project problems can
be attributed to the lack of communications. This can include
the correct definition of the scope and expectations. System
guarantees and requirements usually become the focal point if
a system does not meet performance standards. Documented
communication of project expectations, and acceptance by
engineering firms or contractors, are required to gain solutions
to project disputes.
Document control also extends to plan and specification
review and the expectations of the process. Frequently, systems are designed by a consultant or contractor without a clear
understanding of the review and approval process. The plant
management is asked to review a complicated combination of
engineering controls and equipment. A determination must be
made as to who is qualified to conduct a review of and conduct
system plans and specifications. There must also be communication among the project team, outside resources and plant
personnel regarding other implied approvals. This includes
issues such as consistency between architectural, structural
and mechanical drawings, interferences on drawings missed
and who is responsible for back charges.
2.7
PROJECT TEAM ORGANIZATION, SELECTION
AND SKILLS
Selection of a project manager and support staff depends on
the size and complexity of the project. Smaller jobs may only
require the plant engineer to serve in a part-time role, but complex installations may require full-time leadership and responsibility.
Continuity between the development of concepts and delivery of the final completed project is important. The receiver of
a completed installation should be involved early so expectations are known and operator and maintenance training are
accomplished.
The organization, even on small projects, should be defined
Preliminary Design and Cost Estimation
explicitly. During temporary assignments, there may be role
changes that may not be compatible with normal plant or company organization. Project success requires company management’s support of these role changes. It is important to include
anyone who has the ability to change or delay the project. For
environmental and employee exposure control projects this
would especially apply to health and safety staff.
Personnel not normally familiar with the disciplines and
schedule requirements of a complex installation can be included in projects. Thus, care must be taken to properly train all
project team members. At a minimum this training should
include: 1) cost management; 2) schedule control; and 3) communications skills.
2.8
INTERNAL RESPONSIBILITY FOR FINAL
APPROVAL OF BUDGET, TECHNICAL MERIT AND
REGULATORY ISSUES
After building the project organization the project team
responsibilities are determined. The primary purpose of the project team is to manage the installation. One of the first determinations is under what conditions the installation will be accepted. This acceptance may require more than one set of conditions.
The installation is usually impacted by regulations that can
include: improvements to plant ambient air conditions, safety
requirements, requirement to meet emission regulations, and
installed equipment plant and regulatory safety requirements.
As mentioned earlier, the approval process can include
review of complicated engineering drawings, calculations and
specifications. This can include an implied approval of items
such as physical dimensions or connections to plant equipment. These approvals may have cost impacts. For example, a
plant project team may be asked for the selection between two
alternate control schemes that have cost, technical and regulatory implications. Members can be well versed in plant and
process operations, but may not possess the technical expertise
to approve these issues. At that point, outside resources may
need to be considered.
Ultimately the project manager is responsible to management and must sign off on all decisions (Figure 2-1). A team
member may be designated for review of certain aspects of the
installation, but final approval must come from the project
manager. The important factor is ensuring that the project
manager is not inundated with non-critical details related to the
implementation of the project, such as review of each dimension on a drawing. The project manager’s time should be
devoted to larger problems such as ensuring that the most efficient and effective system is installed. Good communications
between the project manager and team members helps ensure
that decisions are timely and the project remains on schedule.
2.9
COMMUNICATION OF PLANT (AND PROJECT)
REQUIREMENTS
At this point, the project team turns outward to communi-
2-9
cate the project’s requirements to the party responsible for system design and to the user. In simplified terms, a project design
can be considered to have three distinct steps: Conceptual
Design, Design Definition and Detailed Design.
2.9.1 Project Feasibility and Conceptual Design. During
this phase, a plan is developed to define feasibility and preliminary design; the minimum requirements from this part of the
project require the following information:
1) Clearly stated objectives (i.e., exposures below “x”,
reduction of environmental emissions by “y”, etc.)
2) Concept description
3) Equipment list
4) Process Flow Diagram (PFD)
5) Heat and material balance
6) Process and Instrumentation Diagrams (P&IDs)
7) List of existing equipment (e.g., fans, motors)
8) Instrument list
9) Milestone schedule
10) Time critical dates (e.g., scheduled maintenance shutdown, production schedules)
11) Preliminary cost estimate
12) Available utilities
13) Building codes and regulatory requirements, e.g., environmental permits (changes to existing permits,
increase in emissions, etc.), relevant fire codes
14) Equipment layout, floor space, site location
15) Studies list (fire protection, safety, etc.)
16) Proofs of performance for existing systems
17) Any end-user system maintenance and monitoring
training to ensure long-term operational reliability.
The final documents include the compilation of all of the
above information. These documents may be accumulated
from internal and external resources, but should be in a format
suitable for presentation to the owner of the system.
The level of accuracy at this stage must be sufficient to identify major problems impacting the final installation cost. Note
that some option analysis may not be resolved until the project
reaches design definition, but should be defined before proceeding with detailed design.
2.9.2 Design Definition – Defining and Communicating
the Scope. After completion of the conceptual design, the sec-
ond phase is developed. For small systems, this may be done
with a few lines of description. In more complicated projects,
a formal design basis (Figure 2-3) becomes the method of
communication once all studies are complete, and all design
options have been determined. This information can be frozen
in place during long lead time items like obtaining permits and
equipment purchase and installation.
2-10
Industrial Ventilation
FIGURE 2-3. Sample Design Basis form
Preliminary Design and Cost Estimation
The design basis is authored by the project team and is a
detailed set of instructions to the design team. This also
includes a list of the expected deliverables at the final detail
design phase (including project goals). The issuance of the
design basis may include other review requirements where
other parties review the conceptual design before proceeding
with detailed design. This second review is used for more
complicated projects where other company and outside
resources may be needed to refine the concepts. For example,
the project team may:
•
review alternate control schemes, costs and schedule
implications
•
pass this review on to corporate representatives or consultants
These steps take longer in this phase but may actually
reduce overall project schedule and costs, by reducing confusion during the design definition and detailed design phases.
Since the design basis document comes from the project
team, the first decision would be who will author, publish and
review the document. Again, project size may influence the
need for text and standards input. The design basis should be
signed by all team members before submission to the selected
design manager, firm or team. It also becomes the scope definition for competitive design bids if that is the direction taken.
At a minimum, the design basis must include the expectations for the project; any applicable standards that must be met
and proofs of performance. These may include regulatory
requirements: American National Standards Institute (ANSI),
NFPA, plant or local standards, safety, and delivery requirements (drawing methods and detail, schedule, etc.). In addition,
adjacent plant environments should be protected from project
stressors (noise, heat, pollution exposures). A sample design
basis is shown in Figure 2-4. At this point, the owner should
have an idea of the annual energy budget and O&M costs.
The design basis is then given to the chosen design firm or
individual and becomes the document for management of
detailed design. The completed design basis can also be used
as scope instructions to design/build contractors (see Sections
2.10 and 2.12).
The extent of the detailed design is also to be determined
during this issue. For example one owner may want all of the
engineering completed as one package. In addition to the ventilation design there may be a requirement for the design of the
electrical power and control, foundations, structural tie-ins to
the owner’s building, etc. Other companies may have in-house
resources already in place for these services. Or, they may
have an electrical design or contracting firm that does all electrical power, controls and/or energy management in the plant.
In those cases, it may be better to have this work performed
outside the ventilation design contract as long as information
is freely transmitted between all parties.
Similarly, other organizational issues can be determined at
2-11
the issuance of the design basis including the distribution of
drawings and review methods for approval of designs and contractor prints. It can also lay out very specific limits of responsibility. As an example, the requirements of the engineer to
review certified prints from vendors making sure that foundations match anchor bolt layouts of equipment. The more information included in the design basis the less opportunity there
is for dispute and project cost overruns.
2.9.3 Detailed Design. This final phase is the one most
identified with the project. It is the final set of instructions to
the installer. Details of design considerations for all of the
major components of systems are included in Chapters 4, 5, 6,
7, 8, and 9. In addition, the calculating methods for system sizing are included in Chapter 9. This phase also includes the
final review set of plans and specifications that the plant management sees before construction bidding. In the case of
design/build contracts it represents the document deliverables
for the installation.
At a minimum, this phase should include enough detail to
clearly communicate the final system to be installed. Drawings
must be to the detail level requested in design basis and may
have company drafting standards included. Normally the contract would require the completion of as-built drawings, and
the turnover of electronic copies for the plant’s files. The level
of detail may extend from single line drawings with few
dimensions to extensive double line drawings that show
details for shop fabrication. Since the cost differential between
the two can be extensive, it is important that the design basis
communicate the expectations of the project and plant management.
Specifications may be included on the drawings or added as
a separate document per various industry and company standards. There are advantages to both methods. On smaller projects the inclusion of information on the drawings keeps one
single source of information for future reference since specification books may be stored in different locations from the
drawings. Larger and more complicated projects may be better
served by the use of specification text packages that may have
clearer information for transfer to the contractor.
At the completion of detailed design, a more defined construction schedule usually can be determined and should be
included, especially on design/build projects. At the same
time, certified vendor prints and cut sheets should be included
in the package.
2.10
DESIGN/BUILD, IN-HOUSE DESIGN OR OUTSIDE
CONSULTANT
After the project team has begun with the development of
the design basis, a decision should be made regarding the
method of design completion. Each of the three methods listed
above has its advantages and disadvantages and the choice
may reflect the preferences of the plant management and the
project team. This may also define the level of instructions in
2-12
Industrial Ventilation
FIGURE 2-4. Sample Design Basis
Preliminary Design and Cost Estimation
FIGURE 2-4 (Cont.). Sample Design Basis
2-13
2-14
Industrial Ventilation
the design basis.
emissions to the environment.
Detailed information may be required for consultants especially if there is a bid process among firms that may not have
done previous work in the plant. The design basis becomes a
bid document (either for design firms or for design/build firms)
and will also be reviewed by all unsuccessful bidders. It may
contain proprietary process information, security and secrecy
of process and/or planning. Thus, document control is important during early organization especially for the issuance of
secrecy agreements and return of information from unsuccessful bidders.
An important consideration is always cost. The project team
must consider all aspects of cost analysis. An engineering firm
with higher hourly rates may actually cost less if their total
hours are less. Similarly an experienced firm may be able to
provide a design that is less expensive and/or more efficient
even if the engineering costs are higher.
When selecting a method of design completion, the plant
management and project team must be realistic in its management abilities. In-house resources are easier to control with
respect to confidential plant processes, but may not have the
depth of expertise to consider different control methods, new
technology or alternate designs. The project may need the abilities of experienced design/build or consulting firms to ensure
that the project meets all requirements. If in-house staffing is
selected then the project would encounter many of the same
issues listed below for design-construct, including review of
capabilities. It should be noted that, as more organizations are
involved in the process, whether in-house or outside, the need
for information hand-offs and reviews increases.
If in-house design is not selected then the next decision is
between the design-construct method and design/build
(turnkey) method. The former uses a detailed engineering package to convey information to the construction contractor
(builder). Sometimes this is known as “plan and spec.” The
builder may be a general contractor, a specialty contractor
(mechanical or sheet metal firm, for example) or a combination.
Table 2-1 outlines considerations when choosing the design and
construction resources (in-house, design-construct or
design/build (turnkey). A common problem is the use of heating, ventilating and air conditioning (HVAC) companies for
the design of industrial ventilation (IV) exhaust systems. The
requirements for these two system types are different even
though both involve the movement of air as IV is a specialized
subset of HVAC design. An HVAC engineer would normally
be experienced in the design of building mechanical systems,
supply duct systems, chillers and air handlers. They may not
possess the required skills to design an industrial ventilation
system that uses air cleaning devices, material handling fans,
heavy gauge duct and involves issues like minimum transport
velocities and hood design. At the same time, an industrial
ventilation firm may be unable to consider all of the requirements of a complicated air conditioning installation. The fundamental difference between IV and HVAC systems is primarily the function of the system and the pressures at which these
systems operate. Additionally, HVAC systems are typically
focused on human comfort in controlled environments, while
IV systems focus on indoor air quality, or human comfort in
much larger, open spaces. Additionally, IV systems may have
environmental conditions that need to be maintained to reduce
The major consideration should be life cycle costs that
include initial capital costs, but also consider the operating
costs over the life of the system. A low initial cost installation
or design by an inexperienced engineer may burden the plant
with high power and maintenance costs for 20 years or more.
2.11
DESIGN-CONSTRUCT METHOD (SEPARATE
RESPONSIBILITIES FOR ENGINEERING AND
INSTALLATION)
The Design-Construct engineering package would contain
sufficient drawings, specifications, logic drawings and other
materials to convey the requirements of the system and the
physical dimensions to contractors for bidding and installation. The drawings may be stamped as required by the project
or the regulatory agencies. This would require a Professional
Engineer for the design.
2.11.1 Selection of Engineering Firm. If the choice is to
proceed with Design-Construct, then the selection of the engineering firm is obviously the next important issue. If the project team is in place, they may choose to pre-qualify one or
more firms for a presentation of experience and capabilities.
These firms could be specialty companies, or departments in
larger multi-disciplined firms, who design only industrial ventilation systems. If a detailed design basis has been developed
to hand over to engineering bidders, the selection process can
move more easily because the definition of scope is usually
clearer. State and national professional engineering societies
also have guidelines for the selection of firms. They focus on
experience and quality of work. Any firm must be able to provide references on similar projects.
During the selection process for the consulting or in-house
engineer, a realistic schedule must be communicated. This
includes milestone dates (permit application and approval,
beginning and end of engineering, construction, commissioning, etc.).
In addition, all of the disciplines must be determined and
division of responsibility made. For example, a design may
include permit application specialists, construction managers,
civil engineers for foundations, electrical engineers for power
and controls, mechanical engineers for the air-moving systems, chemical engineers for process safety reviews and structural engineers for duct and collector supports and roof loads.
After the screening process for an engineering firm is complete, a decision must be made as to the method of payment.
Very large projects may be paid on a fee basis where the pay-
TABLE 2-1. Project Design and Construction Methods – Impact on Owner
Preliminary Design and Cost Estimation
2-15
2-16
Industrial Ventilation
ment is a percentage of the total construction cost. Projects that
fall into the size range of most industrial ventilation systems
usually would be performed on a fixed price or Time and
Material (T&M) basis.
A fixed price proposal is usually the best method for the
user because it is easier to manage budgets. To ask an engineering firm to bid this way, the design basis and schedule
must be very explicit and there must be clear methods
described for scope changes and their management.
Some firms may have a tendency to bid low while asking
for change orders as they occur. Other firms may build some
factor of safety into the price and seldom ask for scope
changes unless they are significant. When checking references
it is important to ascertain the scope management history of
the firms bidding the project. As mentioned earlier, T&M rates
can be very misleading. A company with lower rates can actually have higher total costs because hours are higher (cost =
hours H rate).
It is especially difficult to choose engineering firms for
blanket order arrangements strictly on rates. If forced to be
most competitive on rates, less experienced personnel may be
substituted or qualifications can be inflated to ‘play the rate
game.’ It is always best to choose engineering firms on qualifications, efficiency and experience and not on initial costs.
The premium paid for less experienced engineering is paid
over the life of the project. Also, if a particular engineer or
group has the knowledge and experience for a project or technology, they may need to be specified by name in the negotiation of the contract.
Important issues with regard to selection of installing contractors and the management of the construction are included
in Chapter 1 of the O&M Manual.
2.12
DESIGN/BUILD (TURNKEY) METHOD – SINGLE
SOURCE OF RESPONSIBILITY
This method can be reflected in many different types of
partnerships and can include:
1) An engineering firm as the prime contractor partnering
with an installer to provide a turnkey project;
2) The installer as prime contractor using their own inhouse design staff;
3) The installer as prime contractor using an engineering
firm or other resources for the design; or
4) Partnerships or joint ventures between design and
installation firms to provide turnkey installations.
As with design-construct there are inherent advantages and
disadvantages. When the owner defines the project, the design
basis can be issued as a set of instructions to the design team.
If this approach is selected then the design basis can be issued
directly to the turnkey bidders. When proposals are returned,
the owner can be assured that everyone is bidding to the same
general scope and project requirements.
However, each design/build contractor would be given flexibility in their presentation to take the contract based on their
own preliminary design. This could include different methods
of hooding, different air volumes, different collection devices,
etc. This puts the premium on the experience and capability of
the design/build contractor. The contractor would be absorbing
the risk of the design and delivery. Thus, they would have to
provide a system with enough surety of design to complete the
project profitably, but not be so safe in their choices that they
make their price too high. Where specific control approaches
(i.e., hood type) or mechanical performance specifications
(i.e., minimum capture velocities) are required by the owner,
they should be clearly stated in the design basis and not left to
the discrepancy of the turnkey contractor’s design team.
2.13
PROJECT TEAM AND SYSTEM EVALUATION
This range of acceptability may have to be evaluated by the
project team for either method. It is similar to the selection of
the correct engineering firm when pre-screening for a designconstruct project. The difference is the turnkey method does
relieve the project team of potential burden of ruling on disputes between designer and installer if there is a system failure
on a design-construct project. In these latter cases, the team
must determine if it was a design or installation flaw (or both)
to assign back charges and move the project to completion.
In design-construct, the engineer would design to the standards in the design basis. In every case it costs very little more
for an engineer to design with enough factor of safety to ensure
they have a successful installation. For example, the costs to
design a system with 20,000 acfm [9.44 am3/s] are marginally
more than that required to design the same system with 10,000
acfm [4.7 am3/s]. Moreover, the 20,000 acfm system would
work with more margin of safety than the smaller one.
Unfortunately, the owner pays for this safety factor for the
remaining life of the project. They pay in higher installation
and operating costs.
A design/build proposal forces the bidder to consider their
own risk for performance and safety factor. Since the bidding
would be competitive they must build their expertise and experience into their price. However, the owner must now make his
purchase decision based on the review of many proposals that
may have varied design parameters. One design build firm
may propose 10,000 acfm [4.7 am3/s] and the other may propose 20,000 acfm [9.44 am3/s] for the same process. The project team must now decide which is correct (and possibly
ignore the price implications). They must also decide if the
company giving the lower price and smaller system can provide the guarantee if there is non-performance. Life cycle cost
analysis by the User will help determine which proposal is
viable in the long term. Although there are instances where
turnkey design/build proposals vary due to different proposed
solutions to a given problem, the owner can also take steps to
Preliminary Design and Cost Estimation
eliminate the degree of variance in those bids by performing a
front end engineering and design study. By completing this
effort, the system operating parameters can be determined
beforehand, eliminating the potential for significant scope differences by turnkey contractors. The front end effort is essentially a conceptual design with potential variable factors in the
final design eliminated; emission points, volumetric determinations, hood concepts, duct sizing, preliminary duct layouts,
type and size of control device, and an estimate of the fan specification can all be specified for the turnkey bidders prior to
bid submission.
2.14
PROJECT RISK AND NON-PERFORMANCE
The implications of non-performance go beyond the obvious. For instance, a system that cannot meet guarantees of
emission levels may delay the start of a process installation.
This delay can have an economic impact that greatly exceeds
the ventilation system’s cost. Similarly, the system may actually work and perform to standards, but may require inordinate
amounts of maintenance and other resources to keep running.
These factors may not have been included in any performance
guarantee.
Using the design-construct method (separate design and
installing firms) opens possibilities for conflicts as the installation progresses. Drawings furnished by the engineer may be
inaccurate or incomplete providing opportunity for the
installer to recover extra costs associated with these errors or
omissions. In the United States Supreme Court case United
States v. Spearin,(2.1) it was determined that the owner warranted that the engineering package (drawings and specifications)
was accurate and sufficient to build the project. (Note that this
Manual is not intended as a law reference and that court rulings can be altered at any time by review and appeal.)
From that case, the Court’s decision produced what is now
called the Spearin Doctrine.(2.1) Under that doctrine, the contractor can recover from the party who supplies the plans and
specifications (usually the owner) the costs for delays and
added costs due to errors or omissions. In response to this doctrine, many owners include provisions in the bidding and contract documents to lessen its effects. These may include requiring the contractor to assume responsibilities for final checks of
the drawings. This can become a complicated issue on large or
risky projects that have the potential for cost and design disputes. Because disputes may eventually happen between consultant and owner, it is important in the selection process to
have agreement on all responsibility issues before design
begins.
The project team must be informed in order to make clear
and concise decisions and may need legal input on complicated
installations. The idea that using design/build methods relieves
the owner of any responsibility is short-sighted. The contractor
may give a guarantee, but the financial strength and quality of
the guarantor must be determined. At the same time, precise
2-17
expectations must be communicated to the contractor. Certain
financial and time aspects may be tied to the performance as
long as these are clearly stated during the bidding process so the
contractors can include these risks in their bids.
It is just as important to make these requirements realistic
and enforceable. Requiring extremely low dust levels may not
be possible because ambient levels from nearby areas may
already be higher than the guarantee request. Issues such as
housekeeping, material handling methods or other factors may
be completely outside of the control and scope of the ventilation system. Even though the owner may get a guarantee from
a consultant or installer, the reality may be that this guarantee
can never be enforced.
2.14.1 Communication of Risk. Any time a contract is
entered into between the owner and outside supplier(s), risks
in the delivery of that contract will exist. The project team
must make an assessment of these risks and determine how
much of the risk should be shared by the other parties. Systems
that have a history of simple and predictable operation may not
have much to consider for risk costs or contingencies.
However, all risks must be considered in systems that are
attached to new processes or involve new technologies. It is
always best to communicate these risks to all parties before
contracts are signed so that a plan is in place in case the systems do not meet the requirements.
These communications include, but are not limited to:
1) What are the expectations of the system at start-up (emissions, in-plant dust levels, bag life, pressure drop, etc.)?
2) What are the expectations of the system during normal
operation (are contingencies necessary for an accidental spill, fire or explosion, i.e., a purge, a full shutoff of
one or both of the supply and exhaust systems)?
3) Does risk free imply excess costs to ensure compliance?
4) What outside influences can affect the guarantee of the
system?
5) How should risk factors be conveyed to vendors (engineers and product suppliers)?
6) Who absorbs risks: engineer? equipment supplier?
contractor?
7) Should design follow or comply with published guidelines or some other recommendations?
8) What procedure is in place to mediate the conflicts
between parties if there is a failure to meet guarantee?
Any system provided by the three methods discussed in
Section 2.10 must have the same goals. They must meet all of
the regulatory, process and safety requirements of the project
at the lowest life-cycle cost. These lowest costs are never really totally known even after the project is installed and running
to specifications. Nevertheless, as the project proceeds and
develops, the project team would be required to use its experience, training, outside resources and judgment to make the
2-18
Industrial Ventilation
best decisions to meet the goals.
2.14.2 Communicating Proof of Performance. Proof of
Performance is the defining requirement for any installation.
This guarantee could be limited to meeting the intent of the
design basis (i.e., provide an air volume of “x” acfm with a
minimum transport velocity of “y” fpm in all duct branches),
meeting other guidelines such as ACGIH® recommendations
for hood design, or meeting all applicable codes and regulations such as in-plant dust levels or emissions. It should be
noted that hood design recommendations provided by
ACGIH®, ASHRAE or similar resources may be stated in a
range (i.e., control velocity of 150 to 250 fpm). If a proof of
performance is based on these references, then it would be
necessary to focus on values within those published ranges.
Note that proof of performance of the project may not protect
the environment and workers from stressors.
It is easiest to demand a proof of performance based on regulatory levels, such as Permissible Exposure Limits (PELs) or
emission limits. It is also important to determine the factors
that can be controlled by the system. For example, if a system
is designed to control worker exposure to dust, it is important
to know background ambient dust levels in the plant before
installation of the system. One method may be to take area
exposure levels in the plant at key locations near the new system. Measurements could be taken before the system operates
and after the system is commissioned. If background exposure
levels before system operation already exceed the guarantee,
the new system may never be able to meet requirements.
See Chapter 2 of the O&M Manual for details of the system
design, installation and project teams when proceeding with the
commissioning process and verifying proof of performance.
2.15
USING PLANT PERSONNEL AS PROJECT
RESOURCES
Ultimately the system is received and used by plant personnel. After final acceptance, there is usually a person designated
as the receiver of the completed project. This may be the plant
manager, operations manager or maintenance manager. In
addition, someone would be designated as responsible for the
ongoing operation and maintenance of the system.
There have been many documented cases where successful
installations meeting all startup guarantees are altered,
removed or even sabotaged after the contractor leaves the site.
This may happen because the installed system does not represent a workable solution to meet the production, access or
maintenance requirements. Whether cardboard is mounted
over hood openings, replacement air systems are diverted,
hoods are removed, alarms are disconnected or controllers or
fans turned off, the results are the same. An expensive system
installed with the best of intentions is left idle or debilitated not
meeting its intended goals.
When the project team includes operations and maintenance
personnel (the end users), project goals are easier to manage
into the commissioning phase. This is because the end users
had input early in the design process and bought into the project. Even for small projects, the experience of the operator
helps ensure that the system will be used and maintained. This
is because the end users had input early in the design process
and bought into the project of those actually using the equipment. This information should be gathered using a questionnaire format or at least through interviews with written comments.
Similarly, maintenance implications can also make or break
an installation. This includes maintenance access to the
process being ventilated, and the ventilation system hardware.
Collectors should be selected for easy access bag removal and
replacement. Fans should be properly fitted with power transmission guarding. Remember the $20 bill test. If design of the
guarding would allow an employee to reach above, below or
through the guard to remove a $20 bill from the pinch point,
the guard is inadequate. Motors and controls should be specified to match existing capabilities and training or additional
operator and maintenance training provided.
Access to duct and equipment should include work platforms and proper ladders or stairs to get materials and equipment to high maintenance areas. Many times a large system
may require additional personnel to address new system maintenance and operational issues. Planning for these needs while
the project is still in the development and installation phase
can save training costs and avoid possible safety issues.
Changes to plant operations may include new requirements for
safety and fall protection.
2.16
INTERFACE BETWEEN THE PLANT AND
PROJECT
The plant must be prepared for major new construction.
This includes on site contractor traffic, plant entry security,
enforcement of fire safety regulations and contractor use of
receiving docks, restrooms and cafeterias. Construction
requirements must be coordinated with production, shipping
and other plant needs to reduce interferences.
Normally the project manager may need to be involved in
securing permits for construction and operation. Some permits
may require long lead times and must be included in project
schedules at the beginning.
The installation of the system would also impact the plant in
other ways. The auxiliary equipment requirements for the system itself is the most obvious. Additionally, the plant’s electrical power, compressed air, water, sewage or other systems
may be inadequate or require intermittent shut-down periods
during phases of the installation process. This should be considered during the ventilation design and suitable plans should
be included for expansion of these systems.
Similarly, any new exhaust system should include consideration for replacement air. If the plant is in balance (supply air
equalled exhaust air) before the project, it may just require a
Preliminary Design and Cost Estimation
supply volume equal to the exhaust. If the plant was not in balance, the project manager may want to use this as an opportunity to cure under-design of replacement air by adding more
air supply to newer projects. In any case, the placement of the
supply air may have effects on adjacent areas not normally
considered in the project.
2.17
IMPACT OF NEW SYSTEMS ON PLANT
OPERATION
There can be unpredicted influences on the operation of the
new system. Because the process itself may now be more
enclosed to provide better contaminant capture and control,
there may be heat build-up. This can translate to higher duct
and system temperatures. It also may cause formations of different chemicals in the exhaust gas streams or change the dew
point or acid dew point.
If the exhaust system now includes long runs of hot duct
there can be condensation issues that had not been accounted
for in design. For systems involving heat and moisture in the
gas stream, it is important that the project team consider these
effects on the plant environment as well as the plant’s effect on
the local exhaust system.
Frequently, other issues may arise when a new local exhaust
system is installed. The local exhaust system must meet its
stated goals but also may cause other issues that must be
addressed during installation.
Final success criteria for the preliminary design phase is the
definition of a project that meets all of the regulatory, safety
and operations needs of the plant. It is then feasible to move to
detailed design phase.
2.18
CAPITAL COSTS
Determination of capital cost is one of the first requirements
in determining the feasibility of a project. For some projects,
the industrial ventilation (IV) system can contribute a significant percentage of the total project cost. There can be several
estimates made during the course of the project, usually with
increasing accuracy. The first (and least accurate) is an “Order
of magnitude” estimate made very early in the project.
2.18.1 Order of Magnitude Pricing Methods. There are
several methods for determining an order of magnitude cost
estimate. None of these methods has a reliable or repeatable
accuracy of less than ± 30%. While inaccurate, these methods
can be an invaluable first step in determining the feasibility of
the project. These methods are also sometimes called first cut
or rough cut pricing. There are two main types of Order of
Magnitude cost estimating.
Rule of Thumb Pricing. This method is based on historical
price data on similar applications. If, for several past projects,
a baseline unit and a ventilation system cost are known, a rule
of thumb can be developed. The baseline can be one of several
known performance indicators of a process. Commonly used
2-19
baselines include: tons of processed material per day, heat output of a process, volumetric flow of exhaust system, or units
per hour of production. As an example, consider a ventilation
system with an estimated volumetric flow. Using several projects, an average cost can be developed using units such as control dollars per acfm ($/acfm). Then the size of the proposed
system can be multiplied by the projected system size to determine the rule of thumb price.
EXAMPLE PROBLEM 2-1 (Rule of Thumb Pricing)
In the evaluation of five dust control projects at a gray iron
foundry in the cleaning department, the average cost for the
exhaust was found to be $15/acfm. A shot blast operation is
being added for a new casting that will require 20,000 acfm of
exhaust. Determine the rule of thumb cost estimate.
When applying a rule of thumb estimate the impact of
expensive auxiliary items should be identified. If the item is a
variable to the system it should be removed from the projected
cost. Examples of such items are electrical substations, steel
reinforcing, significant steel supports or relocation of significant utilities.
Scaled Pricing Factor. If the proposed system is similar to
an existing system where the project cost and system size are
known, the scaled method can be used. The only variable
required to calculate in this estimate is a volume estimate for
the proposed system. The cost can be determined using the
Power Rule. If a baseline cost and the corresponding system
size are known, the cost for the proposed system can be estimated using the equation:
[2.1]
where:
Cb = baseline cost for previous project
Ca = estimated cost for new project
Sb = baseline system size for previous project
Sa = estimated system size for new project
The system size can be expressed in volumetric flow, production output, or any other scalable function associated with
the controlled operation. Estimates found using this method
2-20
Industrial Ventilation
are only recommended if a single project cost for a similar
application is known.
Purchase Price Estimate (A) = $225,000
Purchased Equipment Cost (PEC) = (1.18)(A)
= (1.18)($225,000) = $265,500
Total Capital Investment = (Installation Factor)(PEC)
= (2.19)($265,500) = $581, 400
EXAMPLE PROBLEM 2-2 (Scaled Pricing Factor Example)
(IP Units)
A system had been installed recently with an air volume of
24,000 acfm and an installed cost of $420,000. Included in the
cost of the project was a new substation worth $100,000. A new
similar system is being considered but 32,000 acfm is required.
Estimated Engineering Costs = (0.10)(PEC)
= (0.10)($265,500) = $26,550
NOTE: Many of the cost factors can change due to special
considerations; for example, sales tax varies by region and
state.
Base system cost = Cb = $420,000 – $100,000 = $320,000.
2.18.2 Equipment Factor Pricing Method. Determining a
cost estimate for the system can be made more accurate if the air
control device and system volume are known. It is common for
an industry to use one predominate type of control equipment for
a given application. An example of this is the use of catalytic
oxidizers for the control of volatile organic compounds (VOCs)
from printing operations or baghouses (fabric filters) for foundry
surface treatment operations. For a given control method, the
cost of equipment is multiplied by factors to estimate related
costs. This method is documented in the publication “The EPA
Air Pollution Control Cost Manual”(2.2) and is provided at no cost
from the USEPA website. Estimates using this method can be as
accurate as ± 20% depending on the project.
2.18.3 Price Book Estimating Method. The methods in
this section are more accurate (± 10%) but require considerably more time and effort to complete. The design phase of the
project (see Chapter 2, Section 2.9.3) must be complete before
this method can be used. A price book is a listing of equipment,
fabricated pieces and related installation costs in an indexed
format. There are price books available commercially and
many contractors and fabricators have prepared price books
for their internal use.
The project must be broken down into the individual pieces
and each separate piece must have an individual cost estimate
prepared. The first step to preparing this type of estimate is
called the “take off.” Using the design drawings for the system, each individual component in the system is listed in a
An example of this method is shown in Figure 2-5 and was
taken from the referenced USEPA Cost Manual. The calculation shown is designed to determine the total project cost for
the installation of a fabric filter. The only input to the sheet is
the purchase price of the collection device. This is available
from vendors if a detailed specification can be provided.
Based on that figure, all of the other associated costs can be
estimated as shown. The USEPA has prepared estimating
spreadsheets for most commonly available types of control
equipment, duct systems and exhaust stacks.
EXAMPLE PROBLEM 2-3 (US EPA Cost Manual) (IP Units)
A process requires the use of a fabric filter for control of
emissions from a process. The estimated equipment cost is
$225,000. No building or special site preparation is required.
Determine the estimated total capital investment for the installation and the estimated engineering cost using the values provided in Figure 2-5.
FIGURE 2-5. Equipment factored estimates
Preliminary Design and Cost Estimation
spreadsheet format. Information such as component size,
length of component required, number of components needed
are compiled on the spreadsheet. Using the price book, the cost
of purchasing the item and the installation labor requirements
can be found. If an item is not included in the price book, a
determination of the purchase price and installation labor must
be made. Since some special industrial ventilation components
are not commonly purchased items (special hoods and fittings,
etc.) they must be independently estimated. After the purchase
price for equipment and labor requirements are calculated,
overhead expenses and the desired profit are determined and
added to complete the cost estimate.
Because of the time and expense involved, price book methods are rarely used by owners or designers to determine a cost
estimate. These methods are primarily used by contractors and
installers for determining their project bid.
There are a number of sources that provide cost information
needed to develop a project estimate by the price book method.
Several of these publications are listed in the references section.(2.2–2.7) Use the most current information when preparing an
estimate. Some of the listed sources provide annual updates of
the costs provided. Some of the pricing books have adjustment
factors to modify labor cost according to the location of the
project. If the project is to be designed by an engineering firm,
those fees must be added. The same applies for the use of a
construction manager.
2.19
OPERATING COST METHODS
Initial capital costs are only one component of the total system cost impact. In some cases, annual operating costs can be
higher than the original capital cost of the system. Each industrial ventilation system has specific ongoing costs associated with
the installation. These include electrical power, natural gas, compressed air, waste disposal, maintenance and other items.
These estimates are not only made to predict the economic
impact of the system but also can be used for comparison of
vendors’ proposals as well as comparisons between control
strategies (e.g., baghouse versus scrubber, etc.).
2.19.1 Annual Operating Cost Components. Annual operating costs represent the cash outlay requirements for operating and maintaining the system. These are variable depending
on the maintenance needs for the system and the hours of operation. Costs in this category include: electricity, compressed
air, water, natural gas, waste disposal, operating labor, maintenance labor, supervision, replacement and wear parts, insurance, taxes and overhead.
Very soon after the air control device is selected an energy
audit should be conducted. This audit will include the determination of brake horsepower of motors, compressed air usage,
and other utility requirements. For a thermal or catalytic oxidizer, the fuel usage requirements should also be determined.
Energy Costs. Energy costs for the system can be calculated
2-21
by multiplying the projected energy usage by the cost of acquiring the energy. When determining electrical usage, the primary
expense is the operation of the motors for the system fans.
Other electrical devices included on typical systems are
controls, lighting, rotary locks, heaters, pumps, air compressors, shaker motors, and rectifiers (for electrostatic precipitator
systems). Motor horsepower should be converted to kilowatts.
Other items will typically specify the wattage requirements.
Electricity is sold in kilowatt-hours. Since the equipment electrical usage has been calculated in kilowatts, the number of
kilowatt-hours is simply the hours of operation for the system
times the total power requirement.
Operating Labor and Maintenance Costs. All industrial
ventilation systems will require some operating and maintenance labor. The amount of labor can vary significantly
depending on the type of control equipment selected, the size
and complexity of the system and the type and quantity of
material that is collected. Typical maintenance requirements
for industrial ventilation systems that utilize different types of
control equipment are provided in Table 2-2.(2.3)
Replacement Parts Costs. Some components of the industrial ventilation system will require replacement of worn parts
on a regular basis. These items can have relatively low cost,
such as fan belts, or have major cost impact in the case of
replacement filter media (bags or cartridges). Determining
these costs on an annual basis is done by dividing the cost of
replacement parts by the estimated life in years. Note that
labor required to install these replacements must also be considered.
EXAMPLE PROBLEM 2-4 (Yearly Cost of Media)
A fabric filter collector has polyester media with a useful life of
2.5 years. The cost (bags and labor) to replace the filter media
is $10,000. The resulting annual replacement cost estimate for
the filter media is $10,000/2.5 years or $4,000 per year.
TABLE 2-2. Operating and Maintenance Requirements
Control Device
Fabric filter
Operating Labor Maintenance Labor
(hours per shift) (hours per shift)
2–4
1–2
Electrostatic precipitator
0.5–2
0.5–1
Venturi scrubber
2–8
1–2
Oxidizer
0.5
0.5
Adsorption or absorption
systems
0.5
0.5
2-22
Industrial Ventilation
Waste Disposal Costs. This cost can vary widely depending
on the nature of the material collected. Hazardous waste disposal costs have increased dramatically and this trend is
expected to continue. Estimating waste removal costs is
dependent on the quantity of material collected and is typically
charged by the ton of material disposed.
Waste requirements can be estimated in advance using the
emission factors for the process under control. The USEPA
document AP-42(2.8) lists typical waste concentrations in the
exhaust air for many industry classifications. This document
can be found on the internet at the USEPA website. Using
these emission factors and the volume of air collected, a rough
estimate of the amount of collected waste can be determined.
ly combined into a single factor for cost estimates. The
USEPA uses 4% of the total capital investment. When the
equipment has a short life, the depreciation component may
require adjustment.
2.19.3 Factored Methods for Determining Annual Costs.
The USEPA has created a series of spreadsheets for calculating
direct and indirect annual costs for several types of emissions
control equipment including fabric filters, wet scrubbers, and
thermal oxidizers.(2.2) Figure 2-6 provides an example calculation using the EPA method for a thermal oxidizer system.
2.19.4 Annualizing the Capital Cost of a Project. In addition to the direct and indirect annual costs (electrical power,
compressed air, overhead, etc.) listed in previous sections,
there is also a method to annualize the initial capital investment of the project. This is not depreciation, but a method to
show an added annual effect from the actual purchase of the
system. This effect will consider three main components:
EXAMPLE PROBLEM 2-5 (Dust Catch Disposal) (IP Units)
•
The capital invested for the purchase of the system
A proposed process has an emission factor of 2 pounds per
ton of production. A fabric filter collector is used to remove the
dust from the air. The efficiency is assumed to be almost 100%.
If the production is 4 tons per hour and the period of operation
is 6,000 hours per year, determine the annual waste quantity.
•
The cost of this capital (interest charged to borrow the
money for the purchase); this value is used even if the
system is purchased without borrowing funds – it is the
loss of use of this capital
Waste =
• The expected life of the project (will the system last 5,
10 or 20 years before requiring replacement?, etc.)
This information can be used to compare the annual direct
costs to a value that considers the original purchase. By totaling all of these values the design team can determine the real
cost impact of the installation. This type of analysis is very
2.19.2 Estimation of Total Annual Cost. The total annual
cost of the system includes both the direct costs (components
in Section 2.19.1) and the indirect costs of overhead, depreciation, taxes, insurance and administration.
There are two types of overhead: labor and facility. Labor
overhead includes wages, fringe benefits for the operation and
maintenance staff and includes worker’s compensation, Social
Security and pension fund contributions, vacations and health
insurance. Some of these are fixed costs. Payroll overhead is
traditionally computed as a percentage of the total annual labor
cost.
Facility overhead accounts for expenses not directly related
to the operation and maintenance of the control system,
including: plant security, maintenance of restrooms and break
areas, lighting, and parking. The USEPA uses an estimating
factor of 50%–70% of the wage expenses to calculate the total
of overhead.
Depreciation, taxes, insurance and administration are usual-
FIGURE 2-6. Suggested factor unit cost: thermal fluid-bed
catalyst unit
Preliminary Design and Cost Estimation
important when trying to choose the best of different control
strategies. This method is also used for BACT studies when
assessing whether a control strategy is appropriate or economically feasible.
This basis for the annualized cost is the calculation of a
Capital Recovery Cost Factor (CRCF):
[2.2]
where:
i = annual interest rate
n = capital recovery period (depreciable life
of the system)
This factor is multiplied by the capital invested to purchase
and install the system to obtain the annual cost impact. Note
that all of these costs are in present value dollars and do not
consider inflation or time value of money. This simplified
method can then be used for cost comparisons without adding
these effects. This analysis does not include other annualized
costs such as electric power and maintenance, but only covers
the cost impact of the original project price.
2-23
The most common method to compare competing options is
to consider all costs over the equipment life and bring these
costs back to an equivalent cost in today’s dollars. This is
called the present value analysis of costs. There are factors that
are used to multiply future costs to obtain the present value of
those costs. The single present value factor is used to take a
future cost of “X” dollars and provide the equivalent current
cost of that expenditure. The uniform present value factor
takes an annual recurring cost that does not change and computes the equivalent cost of that expenditure over the years
representing the system’s life in today’s dollars. The uniform
present value factor is used on costs that are routine and will
not change significantly from year to year.
These accounting methods are used to determine the life
cycle cost of a project. Life cycle cost analysis compares the
cost of various alternative actions and identifies the least costly
option by predicting the future costs as well as the current
costs of each choice.(2.9) Considering the impact of inflation, all
costs are expressed in today’s dollars to achieve an understanding of the true relationship between alternative costs or
the payback of an investment. The life cycle cost evaluations
require a thorough understanding of the changing value of currency, and also require an estimation of the inflation influence
on all components in the analysis.
Five types of economic analyses are commonly performed:
•
Simple Payback Period — the total investment savings
in the first year divided by the investment cost. This
analysis is used to evaluate options for changing a current system.
•
Abbreviated Life Cycle Cost Method — the comparison of the annualized cost of competing systems.
•
Life Cycle Payback — the time necessary for the alternative to payback the investment considering increases
in operating and maintenance costs as well as the cost
of capital.
•
Internal Rate of Return — the interest being returned
on the investment over the length of the study.
•
Savings to Investment Ratio — the present worth of the
investment’s savings divided by the present worth of
the investment’s first cost.
EXAMPLE PROBLEM 2-6 (CRCF)
A scrubber system has an installed cost of $250,000. The
system is expected to last 15 years before it is replaced and the
annual interest rate at the time of the project was 6.5%.
Determine the CRCF and the annualized cost of the project.
Annualized Cost = (CRCF) ($250,000) = $26,600 per year
for 15 years
2.20
COST COMPARISON METHODS
Cost comparison methods are required to help decide which
option of competing systems, components, programs, etc. is the
best to select. Although the initial installed cost of a system is
important, it could have a significantly shorter life, a much
higher energy use or a greater operating labor requirement than
an alternative option. The purchaser should be certain that the
initial capital savings will offset the higher operating costs over
the life of the system. It is necessary to add all of these costs
over the life of each option and make an accurate comparison.
All but the first two economic analyses utilize the present
worth of an investment’s savings and costs over a time period.
These analyses are used to compare alternative investments or
the value of making an investment compared to doing nothing.
In doing some of these analyses, one must assume an inflation
rate for the cost of money (called the discount rate).
Since most industrial ventilation systems are not an investment in the monetary sense (unless collected materials can be
sold or returned to the process), the classic Life Cycle calculations are not easily applied. In most cases, the installation of
the system would be selected by lowest cost alternative. The
methods in this section emphasize that total annualized costs
2-24
Industrial Ventilation
should still be considered for alternate vendor and control
method analysis rather than just initial capital costs.(2.10)
2.20.1 Simple Payback Period Method. This is the easiest
evaluation method for comparing changes to an existing system
or systems that would replace an existing system. To perform
these analyses add the annual cost and the cost savings that the
alternative system would provide. This total annual savings
would then be divided into the cost to make the change or install
the alternative system. The result is the number of years
required to pay for implementation of the change. The simple
payback period analysis may not correctly lead to the best solution since the effect of inflation and cost of money are ignored.
$70,000. The baghouse system will have a useful life of 20
years before needing replacement. The baghouse system will
have an annual operating cost of $92,000 for electrical power,
maintenance and all other costs. The scrubber system will
require more horsepower because of higher pressure drop and
annual operating cost was estimated at $105,000. Determine
the best control strategy considering annualized costs.
Scrubber System
Annual Capital Cost =
$26,600
(from Example Problem 6)
Annual Operating Cost = $105,000
Total Annual Cost = $131,600
Baghouse System
EXAMPLE PROBLEM 2-7 (Simple Payback) (IP Units)
A heat recovery unit can be installed in an exhaust system
at a cost of $200,000. This unit will provide an annual heating
energy cost savings of $ 55,000. There also would be an associated increased maintenance cost of $ 5,000 per year. What is
the simple payback of this system modification?
Annualized Capital Cost = (CRCF) ($320,000) = $29,000 per
year for 20 years
Annual Capital Cost =
$29,000
Annual Operating Cost =
$92,000
Total Annual Cost
$121,000
Annual Savings = $55,000 – $5,000 = $50,000 per year
Simple Payback Period = First Cost/Annual Savings
= $200,000/$50,000 per year = 4 years
The simple payback of this project can be compared with the
payback of other projects that compete for corporate funding
and decisions can be made as to which ones to implement.
2.20.2 Abbreviated Life Cycle Cost Method. Rather than
considering the time value of money, a simpler method would
be to keep all factors in present value dollars for comparison.
This does not consider the changing of value of currency and
assumes that the effects of inflation are constant for all components of the analysis. This method does not take into
account, for example, that energy costs may be increasing in
value at a faster rate than the cost of maintenance labor.
From this calculation, the baghouse system would have a
total annual cost of about $10,600 less than the scrubber system over the life of the project. Note that this will not always
favor one type of collection device over another, but does show
that annual operating costs as well as other factors such as
useful life of the equipment can have more impact than just initial project costs.
2.20.3 Life Cycle Payback Method. The life cycle payback
method determines the time necessary to recover the initial
investment to install a system. The present value of future
costs and savings are identified for each year in the future.
These values are then added to the initial investment cost. The
year in which savings repays the final part of the investment is
noted as the time period required to achieve the payback.
To determine the future cost of an item that increases by a
fixed percent, ic, each year, multiply the cost for the previous
year by (1 + ic). Then:
Ck = Ck-1 (1 + ic)
EXAMPLE PROBLEM 2-8 (Abbreviated Life Cycle Cost)
The scrubber system listed in Example Problem 2-6 is being
compared to an alternate control strategy (fabric filter baghouse). The initial cost of the baghouse system is $320,000 so
the scrubber system has a potential savings to the project of
where:
Ck = the cost in the year k
Ck-1 = the cost in the year prior to year k
ic = yearly increase in cost
[2.3]
Preliminary Design and Cost Estimation
In conducting the analysis, the cost of energy (Ek), and rates
of change for energy (ie), the cost of maintenance (Mk), and
rate of change for maintenance (im), the cost of operating labor
(Lk), and rate of change for labor (iL), and the cost of replacement parts (Rk), and the rate of change in replacement parts
(iR), or any other recurring cost can be factored into the calculations. The discount rate or estimated inflation rate can also
be input.
The present value for the cost in year k item is given by:
PVCk = Ck/(1 + i)k
2-25
ues of the energy and maintenance savings for the preceding
year’s cash flow. In this analysis, no tax impact is considered.
For example, in year 1 maintenance costs would be $5,000
(1 + 0.03) = $5,150.
The present value would be $5,150/(1 + 0.039) = $4,947.
Table 2-3 tabulates the present values for the next seven
years.
[2.4]
Table 2-3 shows the results of the calculations for the project. Since the cumulative cash flow changes from negative
(loss) to positive (gain) between the fourth and fifth years, the
payback period for the $200,000 investment was a little over
four years.
where:
PVCk = Present value cost in year k
PVC0 = Cost if purchased in year 0
i = Discount rate (cost of capital)
EXAMPLE PROBLEM 2-9 (Present Value Total Costs)
The heat recovery unit of Example Problem 2-7 can be
installed in an exhaust system at a cost of $200,000. This unit
will provide an annual heating energy cost savings of $55,000.
There also would be an associated increased maintenance
cost of $5,000 per year. The annual increase in maintenance
costs is 3% per year. The discount rate (cost of capital) equals
3.9%. The anticipated energy cost increases were estimated
for the next seven years. The cash flow (present value cumulative) is the sum of the initial investment and the present val-
2.20.4 Life Cycle Costing Considering Taxes. The previous analysis did not take into account the implications of taxes
on the cash flow obtained from the implemented system.
Taxes are paid on profits that a company makes when its
income exceeds its costs. Therefore, when operating costs are
reduced the savings are normally taxable income. In addition,
the investment cost of the installed system can be depreciated
over the system life. This depreciated amount is used to offset
income that is taxable. There are different depreciation tables
for installed equipment. The financial department for the specific plant should provide that information as well as the rate
of taxes they pay.
To help understand how taxes and depreciation affect the life
cycle cost the heat recovery system will be evaluated again.
TABLE 2-3. Life Cycle Payback Analysis
Energy
Year
Discount
Rate
Yearly
Increase
0
Yearly
Costs
Maintenance
Present
Value
Yearly
Increase
$55,000
Yearly
Costs
Present
Value
Net
Cash Flow
Present
Value
Present
Value
$(5,000)
$(200,000)
1
3.9%
2.6%
$56,441
$54,322
3.0%
$(5,150)
$(4,957)
$49,366
$(150,634)
2
3.9%
2.9%
$58,067
$53,789
3.0%
$(5,305)
$(4,914)
$48,875
$(101,759)
3
3.9%
2.2%
$59,367
$52,930
3.0%
$(5,464)
$(4,871)
$48,059
$ (53,700)
4
3.9%
2.4%
$60,774
$52,150
3.0%
$(5,628)
$(4,829)
$47,321
$
5
3.9%
1.7%
$61,826
$51,361
3.0%
$(5,796)
$(4,787)
$46,274
$ 39,895
6
3.9%
1.9%
$63,013
$50,088
3.0%
$(5,970)
$(4,746)
$45,342
$
7
3.9%
2.2%
$64,424
$49,288
3.0%
$(6,149)
$(4,705)
$44,583
$ 129,821
(6,379)
85,237
2-26
Industrial Ventilation
EXAMPLE PROBLEM 2-10 (Present Value with Taxes)
The heat recovery system of Example Problem 2-7 provides a
$55,000 energy savings in the first year of operation and there is
an annual maintenance cost of $5,000. The system has an
installed cost of $200,000, which has a $10,000 per year depreciation if a 20 year straight-line depreciation is used. The resulting
savings are taxed at a rate of 40%, which means 60% of the savings is used in the cash flow summation. What are the life cycle
costs and how long is the payback?
REFERENCES
2.1
United States v. Spearin, 248 U.S. 132 (1918).
2.2
United States Environmental Protection Agency: The
EPA Air Pollution Control Cost Manual, 6th edition,
Publication No. EPA/452/B-02-001, USEPA,
Washington, DC (January 2002).
2.3
Vatavuk, W.: Estimating Costs of Air Pollution Control
(1990).
2.4
Ohio, Office of Air Pollution Control, Engineering
Section, Engineering Guide #46, 1983.
2.5
United States Environmental Protection Agency:
Capital and Operating Costs of Selected Air Pollution
Control Systems, USEPA, 450/3-76-014 (Accessed:
May 1976).
2.6
Craftsman Estimating Guides, at craftsman-book.com
(Accessed: 2018).
2.7
R.S. Means Building Construction Data, at
www.rsmeans.com (Accessed: 2018).
2.8
United States Environmental Protection Agency:
Compilation of Air Pollution Emission Factors:
AP-42, 5th Edition, Volume 1: Stationary Point and
Area Sources, USEPA, Washington, DC (January
1995).
2.9
Department of Energy: Software Program, Building
Life-Cycle Costs, BLCC 5.3-09. (Available at: www.
energy.gov/eere/femp/building-life-cycle-programs,
Accessed: May 2018).
2.10
Mullen, M.E.: Moving Beyond Simple Payback,
Addressing Clients’ Financial Needs. ASHRAE
Journal (June 2005).
TABLE 2-4. Example Problem 2-10, Life Cycle Payback Analysis with Tax Implications
Table 2-4 shows the analysis. From the information provided in
Table 2-4, the system pays for itself during the 7th year. In this
example, the discount factor is 3.9% and the energy and maintenance costs escalate at the same rate as with the other examples.
The cash flow changes from negative to positive between the fifth
and sixth years. The payback period is about 5.5 years.
Chapter 3
PRINCIPLES OF AIRFLOW
NOTE: Equation numbers followed by (IP) are designated for inch-pound system only; equation numbers followed by (SI) are designated
for metric use only. If an equation number bears neither designation, then it applies to both systems of measurement.
3.1
3.2
3.3
3.4
3.5
3.6
INTRODUCTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .3-2
RECORDING NUMERICAL VALUES . . . . . . . . . . . . .3-2
PROPERTIES OF AIR . . . . . . . . . . . . . . . . . . . . . . . . . .3-3
3.3.1 Standard Air . . . . . . . . . . . . . . . . . . . . . . . . . . . .3-3
3.3.2 Airflow Terminology . . . . . . . . . . . . . . . . . . . . .3-4
IDEAL GAS LAW . . . . . . . . . . . . . . . . . . . . . . . . . . . . .3-5
DENSITY FACTOR . . . . . . . . . . . . . . . . . . . . . . . . . . . .3-6
VENTILATION SYSTEM PRESSURES . . . . . . . . . . .3-7
3.6.1 Static, Velocity, and Total Pressures . . . . . . . . .3-7
3.6.2 Understanding Pressure Variations Through
a Simple System . . . . . . . . . . . . . . . . . . . . . . . . .3-8
3.7
3.8
CONSERVATION OF MASS . . . . . . . . . . . . . . . . . . . . .3-9
CONSERVATION OF ENERGY . . . . . . . . . . . . . . . . .3-10
3.8.1 Bernoulli’s Equation . . . . . . . . . . . . . . . . . . . . .3-11
3.8.2 Conservation of Energy for Real Fluids . . . . .3-11
3.8.3 Work Done by the Fan . . . . . . . . . . . . . . . . . . .3-12
3.8.4 Heat Transfer Into the System . . . . . . . . . . . . .3-12
3.9 PSYCHROMETRICS . . . . . . . . . . . . . . . . . . . . . . . . . .3-13
3.9.1 Psychrometric Properties . . . . . . . . . . . . . . . . .3-13
3.9.2 Temperature and Humidity Control . . . . . . . . .3-17
3.10 DEW POINTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .3-21
REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .3-21
____________________________________________________________
Figure 3-1
Figure 3-2
Figure 3-3
Figure 3-4
Figure 3-5
Figure 3-6
Figure 3-7
Figure 3-8
Figure 3-9
Measurement of Barometric Pressure . . . . . . . . . .3-3
Uniform and Non-uniform Airflow Profiles . . . . .3-5
SP, VP, and TP at a Point in a Duct . . . . . . . . . . . .3-7
Measurement of SP, VP, and TP in a
Pressurized Duct . . . . . . . . . . . . . . . . . . . . . . . . . .3-8
Variation of SP, VP, and TP Through a
Simple Ventilation System . . . . . . . . . . . . . . . . . . .3-9
Conservation of Mass at a Duct Junction . . . . . . .3-9
Conservation of Mass Across an Air Heater . . . .3-10
Conservation of Mass for an Air–Water
Vapor Mixture . . . . . . . . . . . . . . . . . . . . . . . . . . .3-11
Interchangeability of VP and SP in a
Ventilation Duct (Bernoulli’s Equation) . . . . . . .3-11
Figure 3-10 Work Done by the Exhaust Fan . . . . . . . . . . . . . .3-12
Figure 3-11 The Psychrometric Chart with Identified
Properties . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .3-14
Figure 3-12 The Psychrometric Chart for Example
Problem 3-3 . . . . . . . . . . . . . . . . . . . . . . . . . . . . .3-16
Figure 3-13 Temperature and Humidity Control Processes
Plotted on a Psychrometric Chart . . . . . . . . . . . .3-17
Figure 3-14 Psychrometric Process of Combining of Two
Airstreams . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .3-18
Figure 3-15 Psychrometric Process of Evaporative
Cooling . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .3-20
____________________________________________________________
Table 3-1
Common Physical Constants . . . . . . . . . . . . . . . . .3-2
Table 3-2
Composition of Dry Air . . . . . . . . . . . . . . . . . . . . .3-4
3-2
3.1
Industrial Ventilation
INTRODUCTION
The importance of clean, uncontaminated air for safeguarding workers in the industrial work environment is well known.
A local exhaust ventilation (LEV) system uses air as the medium to capture concentrations of particulates, gases, vapors,
and mists and convey them away from the workplace. General
ventilation systems utilize air to control undesirable conditions
such as heat, odor, and moisture and replace them with clean
and/or tempered air.
This chapter provides fundamental information, terminology, and basic calculation equations used to design ventilation
systems. The properties of air, the ideal gas law, psychrometric
principles, and the relationships between velocity, static pressure, velocity pressure, and total pressure will also be presented in this chapter.
Inch-Pound (IP) Units
•
Area of duct (square feet, ft2): 3 decimal places, i.e.,
0.196 ft2
•
All total and static pressures ("wg): 1 decimal place,
i.e., 1.4 "wg
•
Velocity pressure ("wg): 2 decimal places, i.e.,
1.37 "wg
•
Velocity (feet per minute, fpm): whole values with no
decimals, i.e., 3,562 fpm
•
Temperatures (F): no decimal places, i.e., 77 F
•
Volumetric flow rate (actual cubic feet per minute,
acfm): whole values with no decimals, i.e., 21,456
acfm
•
Loss factors (dimensionless): 2 decimal places, i.e.,
1.78
•
Length (feet, ft): no decimal places, i.e., 5 ft
Table 3-1 and the inside of the back cover provide the basic
definitions, relationships, and constants required for this and
the remaining chapters in both inch-pound (IP) system and
international system (SI) equivalents. A familiarity with basic
scientific notation and definitions is a minimum requirement
for the design of industrial ventilation systems. Before investigating the design chapters in this Manual, one must be familiar with these basic definitions.
•
Area of duct (square meters; m2): 3 decimal places, i.e.,
0.221 m2
•
All total, static, and velocity pressures (Pascals, Pa):
no decimal places, i.e., 247 Pa
3.2
•
Velocity (meters per second, m/s): 2 decimal places,
i.e., 24.05 m/s
•
Temperatures (Celsius, C): no decimal places, i.e.,
21 C
RECORDING NUMERICAL VALUES
The following rules regarding the recording of decimal
places should be used when recording LEV system data and
specifications:
TABLE 3-1. Common Physical Constants
Metric (SI) Units
Principles of Airflow
•
Volumetric flow rate (actual cubic meters per second,
am3/s): 2 decimal places, i.e., 14.84 am3/s
•
Loss factors (dimensionless): 2 decimal places, i.e.,
1.78
•
Length (meters, m): 1 decimal place, i.e., 4.2 m
3.3
3-3
PROPERTIES OF AIR
3.3.1 Standard Air
TEMPERATURE
Temperature is a thermodynamic property that defines the
energy content of the gas stream. As the energy content
increases (or decreases) the temperature will also increase (or
decrease). When two gas streams are in thermal equilibrium,
they are at the same temperature.
In industrial ventilation system design, the temperature of
standard air is 70 F [21 C]. The relationship between the two
temperature scales is:
Fahrenheit (F) = 32 + 1.8 [Celsius (C)]
[3.1a]
In the IP system, the absolute scale is the Rankine scale:
Rankine (R) = 460 + Fahrenheit (F)
[3.1b]
In the SI system, the absolute scale is the Kelvin scale:
Kelvin (K) = 273 + Celsius (C)
[3.1c]
PRESSURE
If a vertical column of air measuring one square inch were
to be weighed at sea level, the pressure exerted by this column
of air would be 14.7 pounds of force per square inch absolute
(psia) in the IP system, or 101384 Pascals (Pa or N/m2) in the
SI system. This pressure is known as standard atmospheric
pressure (Pa). The actual pressure of the atmosphere is not constant and will vary with the elevation above sea level and
weather conditions. Other common methods to identify standard atmospheric pressure include 1 atmosphere (atm), 407
inches of water gauge or column ("wg), 29.92 inches of mercury ("Hg), and 760 millimeters of mercury (mmHg).
Figure 3-1 depicts a barometer that is sealed on top with its
air space at the top evacuated to measure atmospheric pressure. The pressure exerted on the liquid surface at the top of
the vessel would be:
P = ρliquid (L)
FIGURE 3-1. Measurement of barometric pressure
[3.2]
The measurement of pressure relative to atmospheric pressure is defined as gauge pressure (Pg). Gauge pressure is positive for pressures greater than atmospheric pressure and negative when below atmospheric pressure. Various types of
manometers are used to measure gauge pressures in ventilation systems.
Absolute pressure is represented by the sum of gauge and
atmospheric pressures:
Absolute Pressure (Pabs) = Gauge Pressure (Pg) +
Atmospheric Pressure (Pa)
COMPOSITION
Air is a mechanical mixture of several gases with a molecular weight of approximately 28.9 lbm/lbmol [28.9 gmol].
Standard air that is dry (i.e., contains no moisture) has the volume and molecular weight relationship shown in Table 3-2.
Note that the weight and volume of the dust particles in the
air are ignored in the design of dust collection systems. This is
permissible in typical exhaust ventilation systems when the
concentration of solids is less than 25 grains per dry standard
cubic feet (dscf) of air (gr/ft3) or 0.004 lbm/ft3 [57 g/m3] (i.e.,
about 5% by weight). For high concentrations of solids (> 25
gr/dscf [57 g/dscm]) or significant amounts of gases other than
air, corrections for this effect should be included.
DENSITY
where:
P = pressure, lbf/in2 [Pa]
rliquid = density of fluid in vessel, lbf/in3 [g/cm3]
L = length of fluid column, in [cm]
The density of a substance is represented by its mass per
unit volume. The density of air at standard conditions (rstd) is
0.075 lbm/ft3 [1.25 kg/m3] at sea level, 407 "wg [101384 Pa],
70 F [21 C] dry bulb temperature, and zero moisture.
3-4
Industrial Ventilation
TABLE 3-2. Composition of Dry Air
SPECIFIC VOLUME
The specific volume of the air is the reciprocal of the density (1/ρ). For standard air, the specific volume is 13.35 ft3/lbm
[0.17 m3/kg].
3.3.2 Airflow Terminology
IDEAL FLUID
An ideal fluid (including air) has constant density and no
viscosity. Bernoulli’s equation (see Section 3.8.1) is the basic
energy equation for the movement of air; it is based on an ideal
fluid at steady flow conditions. In practice, there are energy
losses with the airflow as it moves through a ductwork system.
REAL FLUID
Real fluids have a property called viscosity (µ), which is the
shear resistance between adjacent fluid layers. When a fluid
(air) is in motion in a ventilation system, the viscosity accounts
for frictional pressure losses from the shear stresses of the air
against the walls of the duct. In practice, Bernoulli’s equation
may be applied to a real fluid like air by adding an energy loss
term [see Equation 3.24a].
speed of sound) is less than 0.3 (since the density change due
to velocity is about 5% in that case). The speed of flow for a
Mach number of 0.3 is approximately 20,100 feet per minute
(fpm) [102 m/s].
STEADY STATE FLOW
The condition in ducts or other industrial ventilation system
components where the flow rate remains constant with respect
to time. An unsteady flow condition would be one where the
airflow rate pulsates or fluctuates with time.
LAMINAR FLOW
The condition where all components of a gas or air are stable against all disturbances and the streamlines move parallel
to the walls of the duct. Laminar flow occurs at low airflow
rates, typically less than 100 fpm [0.5 m/s] in ducts larger than
6 inches [150 mm] in diameter. The Reynolds number (Re),
which represents the ratio of the momentum forces to the viscous forces of a fluid, is 1160 or below for laminar flow. Places
where laminar airflow may occur in an industrial ventilation
system are through a heat exchanger placed in the duct, or
through the fabric filters of a baghouse.
INCOMPRESSIBLE FLUID
An incompressible fluid is one where the density remains
constant. Water is a liquid that is generally treated as an incompressible fluid. In industrial ventilation applications, a gas such
as air is treated as incompressible when the system pressure is
less than or equal to ±20 "wg [5 kPa]. When pressures exceed
±20 "wg, the gas should be treated as compressible and subject
to density correction for the pressure term (see Equation 3.13).
COMPRESSIBLE FLUIDS
Gases and moisture-laden air, where the density varies significantly during flow conditions, are considered compressible
fluids. Compressible flow is relevant in applications when the
Mach number (i.e., the ratio of the speed of the flow to the
TURBULENT FLOW
Turbulent flow occurs at higher velocity conditions in LEV
systems; where Re equals or exceeds 3000. The velocity components in turbulent flow are chaotic and occur both in perpendicular and the general direction of flow. Because industrial
ventilation systems typically operate at velocities greater than
1000 fpm, the flow in the duct is characterized as being turbulent.
Turbulent flow also accounts for those dynamic energy losses in a duct system associated with hood entry, duct fittings,
and transitions.
Principles of Airflow
TRANSITIONAL FLOW
The flow region between laminar flow and turbulent flow is
defined as transitional flow. When Re is less than 1160, the airflow is always laminar; when greater than 3000, the airflow is
always turbulent.
n = m/MW
3-5
[3.3]
where:
m = mass of gas, lbm [g]
MW = molecular weight, lbm/lbmol [g/gmol]
UNIFORM FLOW
The airflow profile in the duct that occurs when the velocity
vectors are both stable and parallel to the direction of flow (see
Figure 3-2) is defined as uniform flow. Uniform flow can
occur in either laminar or turbulent flow conditions; it represents the optimum flow profile for capturing airborne contaminants when designing local exhaust hoods (see Chapter 6).
NON-UNIFORM FLOW
Non-uniform flow occurs when the airflow velocity vectors
are unstable, chaotic, rotational, or not parallel to the direction
of flow (see Figure 3-2). The formation of eddies in the
airstream is a characteristic of non-uniform flow. These currents can result in the failure of exhaust hoods to properly capture and control air contaminants (see Chapter 6).
3.4
The ideal gas law is used frequently in industrial ventilation
applications to calculate the density of a gas or convert the
mass of a gas to a volume basis and vice versa. The ideal gas
law is expressed as:
PV = nRuT
[3.4]
where:
P =
V =
Ru =
=
pressure, lbf/ft2 [Pa]
volume, ft3 [m3]
Universal gas constant
1545.33 ft-lbf/lbmol-R [8.314 J/gmol-K
{Note that 1 Pa = 1 N/m2 and 1 J = 1 Nm}]
n = moles of gas, lbmol [gmol]
T = temperature, R [K]
IDEAL GAS LAW
The behavior of gases is often expressed by use of the ideal
gas law.(3.1) This gas law is useful for estimating the volume of
any gas.
At standard temperature and pressure, one (1) mole of any
ideal gas occupies 387 ft3 [24.1 L]; this is known as the molar
volume of an ideal gas. The number of moles of any gas equals
the mass of the gas divided by its molecular weight and can be
calculated by the following formula:
Note that temperature must always be expressed in absolute
units (i.e., degrees R or K) when using the ideal gas law.
Upon combining Equations 3.3 and 3.4, the ideal gas law
can be expressed on a mass basis:
PV = mRgT
[3.5]
where:
Rg = specific gas constant = Ru/MW, ft-lbf/lbm-R
[J/g-K]
The density of any gas (i.e., ρ = m/V) can be determined by
ρ = (P/Rg)T
[3.6]
Boyle’s law states that volume and pressure of an ideal gas
vary inversely when its temperature is held constant.
P1V1 = P2V2
[3.7]
Another law addressing the behavior of an ideal gas,
Charles’ law states that the volume and temperature of an ideal
gas vary proportionally at constant pressure:
T1/V1 = T2/V2
[3.8]
Finally, Avogadro’s law states that equal volumes of all
gases at the same temperature and pressure contain the same
number of molecules. Avogadro’s law is expressed as:
Ru = PV/T
FIGURE 3-2. Uniform and non-uniform airflow profiles
[3.9]
3-6
Industrial Ventilation
EXAMPLE PROBLEM 3-1 (Volume Determination of
Standard Air)
What is the volume (V) and density (ρ) for 1 lbmol [1 gmol]
of standard air? Note that standard air has a barometric pressure of 14.7 psia [101384 Pa] and temperature of 70 F [21 C].
Solution:
where:
dfe = elevation density factor
dfp = pressure density factor
dft = temperature density factor
dfm = moisture density factor
Elevation Density Factor (dfe)
Using the ideal gas law on a molar basis:
dfe = [1 – ((6.73)(10-6)(z))]5.258
PV = nRuT or
[dfe =
[1 – ((22.08)(10-6)(z))]5.258]
[3.12] IP
[3.12] SI
V = (nRuT)/P
= (1 lbmol)(1545.33 ft-lbf/lbmol-R)
(70 + 460 R)/(14.7 lbf/in2)(144 in2/ft2)
where:
z = elevation of the system, ft [m] above sea level (ASL)
= 387 ft3
Density (r) = m/V = (n)(MW)/(V) = (1 lbmol)(28.948
lbm/lbmol)/(387 ft3) = 0.075 lb/ft3 [SI Units]
Duct Pressure Density Factor (dfp)
V = (1 gmol)(8.314 J/gmol-K)(21 + 273 K)/(101384 Pa)
The pressure of standard air is 407 "wg [101384 Pa].
= 0.024 m3 (Note: 1 m3 = 1000 L; 1 Pa = 1 N/m2)
dfp = [407 + (SPduct)]/407
[3.13] IP
= 24.1 L
[dfp = [101384 + (SPduct)/101384
[3.13] SI
Density (r) = [(1 gmol)(28.948 g/gmol)/(24.1 L)]
(1000 L/m3)(1 kg/1000 g)
where:
SPduct = pressure of the airstream in "wg [Pa]
= 1.201 kg/m3
Temperature Density Factor (dft)
3.5
The temperature of standard air is 70 F [21 C], which is
equivalent to 530 R [294 K].
DENSITY FACTOR
Standard air conditions are seldom achieved in LEV system
design and the cumulative effects of small deviations in air
properties can cause a significant change in density. The calculation of the air density is essential in the design of local
exhaust ventilation systems. The frictional resistance and
dynamic pressure losses calculated for a duct system are proportional to the density of the gas stream. When the airstream
is moist air and is composed of no other ideal gases, we introduce a technique of calculating density using the density factor. Density factor is defined by the following equation:
df = ρact/ρstd
[3.10]
where:
df = density factor (a unitless value)
ρact = density based on the actual conditions in lbm/ft3
[kg/m3]
ρstd = density at standard conditions = 0.075 lbm/ft3
[1.2 kg/m3]
The density factor for gases moving through a ventilation
system is comprised of four components:
df = (dfe)(dfp)(dft)(dfm)
[3.11]
dft = 530/(T + 460)
[3.14] IP
[dft = 294/(T + 273)]
[3.14] SI
where:
T = dry bulb temperature in F [C]
Moisture Content Density Factor (dfm)
Standard air is assumed to be dry air (i.e., contains no moisture). When air gathers moisture, its density will decrease
because water is less dense than air.
dfm = (1 + ω)/(1 + 1.607ω)
[3.15]
where:
ω = humidity ratio; mass of moisture per pound
of dry air (lbmH2O/lbmda) [kgH2O/kgda]
Air is assumed to be at standard conditions (i.e., df = 1.00)
at any point in the design of an industrial ventilation system
when the change in the density (or density factor, df) of the
airstream is less than ± 5% of its value at standard conditions.
Each individual density factor component will attain this condition when:
Principles of Airflow
3-7
1. Elevation is less than 1440 ft ASL [440 m ASL].
3.6
2. Pressure is less than ±20 "wg [5000 Pa].
3.6.1 Static, Velocity, and Total Pressures. There are
three pressures (static, velocity, and total pressure) associated
with an airstream moving through a ventilation system. The
orientation of these pressures for air flowing in a duct is shown
in Figure 3-3. The measurement and prediction of these pressures, and airflow, are the basis for design of industrial ventilation systems.
3. Temperature of the airstream is between 45 F [7 C] and
100 F [21 C].
4. The airstream dew point is less than 80 F [27 C] (i.e.,
 < 0.02 lbm H2O/lbm da [kg H2O/kg da]). The dew
point represents the temperature at which air is 100%
saturated with moisture.
Note that a 3% differential in just two of the four components
of the density factor would result in a 6% net differential in the
overall density factor of the airstream (df); this is significant.
That is:
df = (0.97)(0.97) = 0.94
A 5% differential in all four density factor components would
result in df = 0.81. As such, one must carefully consider the
impact of the airstream density in each industrial ventilation
system design segment; even when each density factor component falls within the specifications noted above.
EXAMPLE PROBLEM 3-2 (Density Factor and Density)
Assume dry air is flowing in a duct at a temperature of 300 F
(148.9 C). The static pressure in the duct is -15 "wg (-3737
Pa), and the ventilation system is in Denver, Colorado, at an
elevation of 5,280 ft ASL [1610 m ASL]. Estimate the density
factor and density of the airstream.
Solution:
Elevation Density Factor (dfe)
dfe = [1 – ((6.73)(10-6)(5,280))]5.258 = 0.83
[dfe = [1 – ((22.08)(10-6)(1610))]5.258 = 0.83]
Pressure Density Factor (dfp)
VENTILATION SYSTEM PRESSURES
Static Pressure (SP) is defined as acting in all directions; it
tends to burst or collapse the duct. It is also the pressure (or
flow work energy) in a duct system that is used to overcome
resistance to airflow caused by duct friction, dynamic losses in
the system (i.e., turbulence losses in duct fittings), and other
losses between the hood and the exhaust stack (such as those
caused by an air pollution control device). Static pressure does
not vary laterally across a duct but does decrease in the direction of flow in a duct with constant diameter.
In industrial ventilation system design, SP is measured with
a manometer; usually in units of inches water gauge or column
("wg) in the IP system, or Pascals [Pa] in the SI system. Static
pressure can be positive or negative with respect to the local
atmospheric pressure and must be measured perpendicular to
the airflow. Holes in the side of a Pitot tube (see Appendix C)
or a small hole carefully drilled into the side of a duct are used
to determine SP.
Velocity Pressure (VP) represents the kinetic energy
required to accelerate an airstream from rest to its existing
velocity. Acting only in the direction of flow, VP is defined as:
VP =
ρV2
2gc
[3.16]
where:
VP = velocity pressure, "wg [Pa]
ρ = density, lbm/ft3 [kg/m3]
V = velocity, fpm [m/s]
gc = mass-force conversion factor, 32.174 lbm-ft/lbf-s2 [1]
dfp = [407 + (-15)]/407 = 0.96
[dfp = [101384 + (-3737)/101384 = 0.96]
Temperature Density Factor (dft)
dft = 530/(300 + 460) = 0.70
[dft = 294/(148.9 + 273) = 0.70]
Moisture Content Density Factor (dfm)
Air is dry, so ω = 0
dfm = (1 + 0)/(1 + 1.607(0)) = 1.00
Density Factor of Air
df = (dfe)(dfp)(dft)(dfm) = (0.83)(0.96)(0.70)(1.00) = 0.56
Density of Air
ρact = (ρstd)(df) = (0.075)(0.56) = 0.042 lbm/ft3
[ρact = (1.201)(0.56) = 0.67 kg/m3]
FIGURE 3-3. SP, VP, and TP at a point in a duct
3-8
Industrial Ventilation
In the IP system, conversions for ft2-to-in2, min2-to-s2, and
psi-to-"wg must be applied; no conversions are necessary in
the SI system. After applying any necessary conversions, and
considering density factor, the formula can be rewritten in the
following form:
nitude in an open duct system is across the fan due to the external energy input. Figure 3-4 depicts the measurement of SP,
VP, and TP in a duct. See Appendix C for a discussion regarding the measurement of pressures in ventilation systems.
[3.17a] IP
strates application of the system design principles using SP,
VP, and TP. The normally vertical exhaust stack is drawn horizontally to show the change in pressures. In the example, the
grinder wheel hood requires 300 acfm (Q) [0.14 am3/s], and
the duct diameter (d) is constant at 3.5 inches [89 mm]. This
diameter is equivalent to an area of 0.067 ft2 [0.006 m2] yielding a duct velocity of 4,491 fpm [22.81 m/s] and a VP of 1.26
"wg [313 Pa] (see Equation 3.17a). The method for calculating
these values is presented in Chapter 9.
[3.17a] SI
When solving for velocity (when VP is known), the formula
can be algebraically rearranged to:
[3.17b] IP
[3.17b] SI
where:
V = velocity, fpm [m/s]
VP = velocity pressure, "wg [Pa]
df = density factor (dimensionless)
Because VP is a measurement of the kinetic energy in an
airstream, and as evidenced by these equations, it cannot be
negative. Note that VP cannot be measured directly; both TP
and SP are measured and then SP is subtracted from TP to
yield VP.
Total Pressure (TP) is defined as the sum of SP and VP:
TP = SP + VP
[3.18]
Total pressure can be positive or negative with respect to
atmospheric pressure; it represents the total energy content of
the airstream, always dropping as the airflow proceeds downstream through a duct. Air or any other fluid will always flow
from a region of higher to lower TP in the absence of the addition of work (i.e., a fan). The only place it will increase in mag-
3.6.2 Understanding Pressure Variations Through a
Simple System. Analysis of the system in Figure 3-5 demon-
In this example, the graphical relationship among TP, VP,
and SP is maintained per Equation 3.18. All pressures are zero
some distance from the face of the hood. Inducing airflow into
the face of the hood requires work by the fan. The SP loss of
the hood is the combination of the pressure energy, or turbulence loss, due to the shape of the hood plus the change of the
kinetic energy of the air to accelerate it from rest to the velocity achieved in the duct.
A grinder hood with tapered takeoff has a SPh = -1.76 "wg
[-438 Pa] and is shown at Point 2 on the static pressure plot. VP
was already calculated as +1.26 "wg [313 Pa] so TP at Point 2 is
calculated as (-1.76 + 1.26) = -0.5 "wg [-438 + 313 = -125 Pa].
As the air and particulate proceed toward the fan, additional
friction and static pressure losses are accumulated. This is
shown on the static pressure graph as the slanting line ending
at Point 3. The friction and static pressure losses from Point 2
to Point 3 total 1.5 "wg, and the static pressure at Point 3 equals
-3.26 "wg: SP3 = -1.76  1.5 = -3.26 "wg. Velocity pressure is
constant so there is a corresponding change in TP for this segment. The TP at Point 3 is SP3 + VP3 = -3.26 + 1.26 = -2.0 "wg.
FIGURE 3-4. Measurement of SP, VP, and TP in a pressurized duct
Principles of Airflow
3ṁ in = 3ṁ out
3-9
[3.19]
where:
ṁ = mass flow rate, lbm/min [kg/s]
The mass flow rate can also be written as:
ṁ = ρQ = ρVA
[3.20]
where:
ρ = density, lbm/ft3 [kg/m3]
V = velocity, fpm [m/s]
A = area, ft2 [m2]
This is a general principle in that it contains no physical
constants and, hence, is equally valid for all fluids (air, water
vapor, gas, etc.).
In Figure 3-6, two airstreams flow into a branch entry and a
single flow exits. Using the definition for the conservation of
mass and then rewriting the equation in terms of air density,
velocity, and area:
ṁ1 + ṁ2 = ṁ3
ρ1V1A1 + ρ2V2A2 = ρ3V3A3
FIGURE 3-5. Variation of SP, VP, and TP through a simple
ventilation system
For standard air:
ρ1 = ρ2 = ρ3 = ρstd
V1A1 + V2A2 = V3A3
There is similar resistance encountered in the straight duct
leaving the fan (Segment 4-5). The static pressure requirements for this segment would also be calculated using
Equation 3.18. Velocity pressure is constant at the outlet of the
fan and is VP4 = 1.26 "wg. The static pressure loss in Segment
4-5 is 1 "wg due to the static pressure resistance of the ductwork. Therefore, SP4 = 1 "wg and TP4 = 1 + 1.26 = 2.26 "wg.
At the outlet of the duct section extending the fan, SP5 is 0
"wg, VP5 is 1.26 "wg, and TP5 = 1.26 "wg.
Finally, the work required by the fan is calculated with
Equation 3.25 or 3.26. Knowing the:
•
Volume flow rate (Q),
•
Fan efficiency from the manufacturer, and
•
The difference between the negative value for TP (or
SP) at the fan inlet and the positive value at the outlet,
Additionally, the volumetric flow rate in any system is defined
as:
Q = VA
where:
Q = volumetric flow rate, ft3/min [m3/s]
Therefore, the classical form of the continuity equation for
air systems is derived:
Q1 + Q2 = Q3
the work can be determined. Chapter 7 details fan energy and
power requirements for fan system installations.
3.7
CONSERVATION OF MASS
In industrial ventilation, the conservation of mass states that
the rate of mass flow into a duct segment by all flow streams
equals the rate at which mass leaves the duct segment by all
flow streams. For steady flow, this can be written as:
[3.21]
FIGURE 3-6. Conservation of mass at a duct junction
3-10
Industrial Ventilation
Figure 3-7 depicts the conservation of mass as applied
across a heater. In this case, there is a change in density as the
air is heated. While the mass rate of flow of air flowing both
into and out of the heater remains identical, the volumetric
flow rate (Q) will change. In this case, the velocity at Point 2
will increase. Thus:
ṁ1 = ṁ2
If the air entering the heater is assumed to be standard air
which is then heated to a new condition with a lower density
(ρ2), then the equation can be stated as:
The exiting stream will have a new volume (Qact) and a new
density (ract). Using Equation 3.20 and knowing the Astd
equals Atotal yields:
ṁtotal = (Qact)(ract) = (Qstd)(rstd)(1 + ω)
Therefore:
Qact = (Qstd)(rstd/ract)(1 + ω)
or:
Qact = (Qstd)(1/df)(1 + ω)
[3.22a]
ρstdV1A1 = ρ2V2A2
Applying Equation 3.21, then:
Equation 3.22a identifies the relationship between standard
and nonstandard air; it is also included as Equation 3 on the
calculation sheet for tabulating pressure losses in a ductwork
system (see Chapter 9). Equation 3.22a can be rearranged to
solve for Qstd:
where:
ρact
df ≡ ____
ρstd
Note that this shows the relationship between standard and
actual air conditions when the density is known. However, it
does not consider the mass of moisture when the airstream
contains water vapor.
Figure 3-8 depicts a system with both standard air (ṁstd) and
moisture (ṁH2O) combining and the mixture of the two leaving. Applying the conservation of mass, factoring out ṁstd, and
then substituting in the term :
(ṁtotal) = (ṁstd) + (ṁH2O)
(ṁtotal) = (ṁstd)[1 + (ω)]
where:
ω = pounds mass of water vapor per pounds mass
of dry air, lbmH2O/lbmda [kgH2O/kgda]
FIGURE 3-7. Conservation of mass across an air heater
Qstd = (Qact)(df)/(1 + ω)
3.8
[3.22b]
CONSERVATION OF ENERGY
Conservation of energy in a ventilation system is the basis
for the equations and formulae used to calculate pressure losses in duct sections. It is also used to determine the work
required by the fan to move the air in a system and is governed
by the first law of thermodynamics. While these principles
have been simplified for application to industrial ventilation
system design, they still guide the overall procedure involved
in system design.
The law of conservation of energy is based on the principle
that energy is neither created nor destroyed. The conservation
of energy considers both mechanical and thermal forms of
energy. In contrast to applications of conservation of mass, not
only can energy be transferred into or out of the system by its
airstreams, but also by non-flow means such as by thermal
input (i.e., heat source or heat exchanger) or mechanical input
(i.e., work provided by a fan).
The equation for conservation of energy in an industrial
ventilation system can be simply written as:
Principles of Airflow
3-11
Again, understanding that the mass flow rate is constant in
this system and applying the definitions inside the back cover
of this Manual, then:
VP1 + SP1 = VP2 + SP2
FIGURE 3-8. Conservation of mass for an air–water vapor
mixture
∑("Energy" In) = ∑("Energy" Out)
∑(ṁe)in + qin + wfan = ∑(ṁe)out
where:
e ≡ (Potential Energy) + (Kinetic Energy) +
(Internal Energy) + (Flow Work Energy)
The “energy” terms are represented by:
Flow Work Energy (SP): SP
ρ
Kinetic Energy (VP): V2
2gc
Internal Energy: u
Potential Energy: gz
Thermal (Heat) Energy: qin
Mechanical (Work) Energy: wfan
Unless the density of the gas being moved through the industrial ventilation system is different than that of air, the inlet and
outlet potential energy terms are equal and can be disregarded.
Therefore, the conservation of energy equation applicable to
the design of industrial ventilation systems can be stated as:
[3.23]
Equation 3.23 is known as Bernoulli’s equation. It is the
basic energy equation for a frictionless, incompressible fluid.
It states that the total energy content of the fluid flowing in the
duct consists of its VP energy and SP energy. For an ideal
fluid, the VP and SP energy terms in Bernoulli’s equation are
mutually convertible (i.e., VP energy can be converted into SP
energy and vice versa).
Figure 3-9 shows air flowing from Point 1, through a contraction to Point 2, and then through an expansion to Point 3.
As the air flows through the converging portion from Point 1
to Point 2, velocity (and VP) will increase as the air flows
through the smaller area at Point 2. This increase in VP is
obtained due to the decrease in SP of the airstream (i.e., an
energy exchange). The opposite occurs as the airflow continues from Point 2 through the expansion section to Point 3
where the velocity (and VP) decreases and the SP increases.
Note that in a real fluid, the conversion of VP to SP or vice
versa is not 100% efficient; because the air has viscosity, pressure losses mainly due to turbulence will be realized.
3.8.2 Conservation of Energy for Real Fluids. Now consider an airstream that is a real fluid possessing viscosity. In
ventilation systems, air flows in a duct from Point 1 to Point 2
and meets resistance. The resistance is in the form of friction
and turbulence that results in pressure losses (i.e., dynamic
losses). Heat is added internally in the air in an irreversible
process resulting from the dissipation of mechanical energy in
the airstream into internal heat. It can be shown that the duct
losses approximate a throttling process wherein the temperature remains constant. In industrial ventilation ductwork systems, the temperature of the airstream will change only when
it passes through a heat or cooling coil, an uninsulated duct, or
a fan.
The conservation of energy equation for a real fluid
becomes:
3.8.1 Bernoulli’s Equation. Consider an application where
the airstream is an ideal fluid (i.e., one that has no viscosity
and has constant density; u = 0 and r1 = r2) flow through a
duct from Point 1 to Point 2. Now assume that there is no
mechanical or thermal energy in this system (i.e., qin = wfan =
0). The conservation of energy equation for this application
becomes:
FIGURE 3-9. Interchangeability of VP and SP in a ventilation
duct (Bernoulli’s Equation)
3-12
Industrial Ventilation
Understanding that the mass flow rate is constant in this system and applying the definitions inside the back cover of this
Manual, then:
SP1 + VP1 = SP2 + VP2 + ρ(u2 – u1)
or:
SP1 + VP1 = SP2 + VP2 + ∑losses1–2
[3.24a]
where:
losses1–2 = energy (i.e., pressure) losses that occur in the
duct system
Pressure losses in a duct system may be caused by:
The energy rise to the airstream occurs due to the TP
increase across the fan and internal energy gain in the
airstream. Note the rise in internal energy is related to inefficiencies in the fan TP increase process. Since there is no way
to evaluate the actual fan work directly, it is determined by
using an efficiency value (h):
[3.25]
When the fan has equal inlet and outlet areas (i.e., the inlet
VP is equivalent to the outlet VP), the work done by the fan is
stated as:
•
hood configuration,
•
duct wall friction,
•
elbows (i.e., turning of the air),
•
branch entry fittings (i.e., turning of the air in the combining streams), and
•
contractions (air is squeezed through a smaller duct or
opening) and expansions.
Additionally, there are other losses in the system such as those
encountered going through air pollution control devices,
dampers, and other assorted fittings.
Equation 3-24a represents the conservation of energy equation for a local exhaust duct system. Using the relationship that
total pressure (TP) is equal to VP plus the SP:
TP1 = TP2 + 3losses1–2
ing air through the duct (see Figure 3-10). It consists of the
shaft energy imparted by the fan energy source (i.e., the motor)
to provide the energy for air movement. It is the primary point
within an air system where the airstream energy level increases (except when there are heating or cooling sources in the system as described in Section 3.8.4).
[3.24b]
Equation 3.24b represents the fundamental concept that, in
any duct section without a fan, the total pressure decreases in
a ductwork system in the direction of airflow.
3.8.3 Work Done By the Fan. Assume a duct section contains a fan to perform work on the airstream. The conservation
of energy equation for this application becomes:
[3.26]
3.8.4 Heat Transfer Into the System. The term (qin) represents the heat transfer into the system between Points 1 and
2 in a ventilation system. When no work is performed by a fan,
and the kinetic energy of the fluid is equal at the inlet and outlet, the conservation of energy equation can be written as:
The heat energy (qin) may also be written as:
qin = ṁh1–2
Therefore:
Using the thermodynamic property of enthalpy:
h=
SP
+ u
ρ
where:
h = enthalpy, BTU/lbm [J/kg]
Simplifying to calculate the work and losses in the fan by
combining values for VP and those shown on the inside back
cover yields:
wfan = Q[(SP + VP)2 – (SP + VP)1] + ∑losses1–2
or:
wfan = Q[ΔTP] + ∑losses1–2
The term wfan represents the work done by the fan in mov-
FIGURE 3-10. Work done by the exhaust fan
Principles of Airflow
For an ideal gas with a mass flow rate in lbm/hr [kg/sec], the
heat transfer rate is determined by:
qin = ṁ Cp (T2 – T1)(60 min/hr)
= ṁ Cp ΔT(60 min/hr)
[3.27]
where:
lations involving air state changes. To determine a state point
on a psychrometric chart, one must know the barometric pressure and two other independent properties:
•
Dry bulb temperature, and
•
Any property that tells us how much water vapor is in
the air.
ṁ = mass flow rate, lbm/min [kg/sec]
A description of the properties on the psychrometric chart is
presented below. The location of the properties is displayed on
the psychrometric chart shown in Figure 3-11.
Cp = specific heat of ideal gas at constant pressure,
Btu/lbm-F [kJ/kg-K]
BAROMETRIC PRESSURE
qin = heat transfer, BTU/hr [kJ/sec]
ΔT = gas stream temperature change, F [C]
Note that the average specific heat (Cp(avg)) can be determined
by:
Cp(avg) = (ṁ1Cp(1) + ṁ2Cp(2) + … + ṁnCp(n))/
(ṁ1 + ṁ2 + … + ṁn)
3.9
3-13
[3.28]
PSYCHROMETRICS
Psychrometrics is the field of engineering concerned with
the physical and thermodynamic properties of moist air and
the processes in which the temperature or water content
change. Moist air is a combination of dry air and water vapor
in varying amounts. Dry air discussed in Section 3.3 is a combination of components in the gaseous phase at the temperature and pressure conditions found in ventilation applications.
However, the water vapor in the air can condense or evaporate
at the same temperature and pressure conditions encountered
in common ventilation systems.
The purpose of psychrometrics is to understand what is happening with the water vapor component as it proceeds through
phase changes with changing temperatures and pressures. In
this section, dry air (da) will be referred to as the air component
and the moisture (H2O) as the water component of moist air.
3.9.1 Psychrometric Properties. The state of moisture-
laden air is defined by its physical properties. These properties
are described below and are depicted on a psychrometric chart
at a single pressure, typically at one (1) atmosphere (see Figure
3-11). There are many types of psychrometric charts in IP
units, SI units, various elevations; some are made for low temperatures and others for medium and high temperatures. An
assortment of psychrometric charts are presented in Chapter 9.
The moisture-laden air is located as a state point on the psychrometric chart for a volume of air. The state point and location on the chart changes as the air volume flows through the
various components of the ventilation system. Depending on
the process situation, the air may be heated, cooled, humidified, or dehumidified. The change in air state can be calculated
with the help of the ideal gas laws presented in Section 3.4.
The psychrometric chart can be used to simplify the calcu-
The barometric pressure labeled on a psychrometric chart is
for a specific barometric pressure. Most charts are based at sea
level conditions where the barometric pressure is 14.7 psia
[101384 Pa]. Psychrometric charts provided at sea level are
valid up to 1,000 feet above sea level (ASL). Above this elevation, the chart needs to be altered because the lower pressure
at higher altitude affects the values of some of the psychrometric properties. If the correct chart is not available, the values of
the other psychrometric properties can be calculated using the
ideal gas law relationships presented in Section 3.4. Another
option is to download or purchase a psychrometric calculator
from the Internet that can calculate the complete list of psychrometric properties at various elevations.
The barometric pressure for a combination of dry air and
water vapor is given by the following relationship:
Pbar = Pda + PH2O
[3.29]
where:
Pbar = barometric pressure
Pda = partial pressure of the dry air component
PH2O = partial pressure of the water vapor component
DRY-BULB TEMPERATURE (T OR Tdb)
Dry-bulb temperature is the heat state of the airstream
observed with an ordinary thermometer. Expressed in degrees
Fahrenheit (F) [C], it may be read directly on the psychrometric chart and is indicated on the bottom horizontal scale (see
Figure 3-11).
WET-BULB TEMPERATURE (Twb)
Wet-bulb temperature is the temperature at which liquid
water, by evaporating into air, can bring the air to saturation
adiabatically (i.e., no heat transfer) at the same temperature.
Expressed in degrees Fahrenheit (F) [C], it can be measured by
wrapping a wet material around the bulb of a thermometer and
moving the air across the wet material.
On a psychrometric chart, Twb is read at the intersection of
the constant enthalpy line with the 100% wet bulb curve line
(see Figure 3-11). There is one point on the psychrometric
3-14
Industrial Ventilation
FIGURE 3-11. The psychrometric chart with identified properties
chart where the wet bulb and dry bulb temperatures are equal;
this occurs where the equivalent dry-bulb and wet-bulb temperatures intersect on the 100% saturation curve. At any point
along the saturation curve, the volume of air cannot hold any
more water vapor molecules without some of them condensing into liquid water.
[1.204 kg/m3]. Lines representing density factor typically do
not appear on low-temperature psychrometric charts when relative humidity curves are presented. To determine the density
factor of moisture-laden air, take the inverse of the humid volume identified at the point of interest on the psychrometric
chart; then divide the resultant by the density of standard air to
obtain the df.
DEW POINT TEMPERATURE (Tdp)
Dew point temperature is that temperature at which the air
in an air-vapor combination becomes saturated with water
vapor; any further reduction in dry-bulb temperature causes
the water vapor to condense or deposit as drops of water.
Expressed in degrees Fahrenheit [C], it is read directly at the
intersection of the saturation curve with a horizontal line representing constant specific humidity (ω), or pound of water
(moisture) per pound of dry air.
RELATIVE HUMIDITY (RH)
Relative humidity refers to the ratio of the actual partial pressure of water vapor in air to its saturation pressure corresponding to the dry bulb temperature. Relative humidity curves
sweep upward from the bottom left on the psychrometric chart
(see Figure 3-11). Values for RH on this curve are determined
by:
RH = 100(PH2O)/(PH2O-Sat)
[3.30]
DENSITY FACTOR (df)
Density factor, as discussed in Section 3.5, is a dimensionless quantity that expresses the ratio of the actual density of the
moisture-laden air to the density of standard air (0.075 lbm/ft3)
where:
PH2O = partial pressure of the water vapor at the dry
bulb temperature of the moisture-laden air
Principles of Airflow
PH2O-Sat = saturation pressure of the water vapor at the
corresponding dry bulb temperature
3-15
ferent moisture contents as well as temperature or other factors
involving density (see Section 3.9.2 for calculation example).
HUMIDITY RATIO ()
HUMID VOLUME (HV)
Humidity ratio represents the pounds moisture (i.e., water
vapor) per pound of dry air (lbmH2O/lbmda) [kgH2O/kgda]; it
may also be expressed as grains of moisture per pound of dry
air (grH2O/lbda)[grH2O/kgda], where 7,000 grains equals one
pound. Also known as the moisture content or mixing ratio of
an airstream, it is used in the determination of the density factor for moisture (see Section 3.5), and in energy calculation for
product drying operations. Humidity ratio is commonly located on the right vertical axis of a psychrometric chart (see
Figure 3-11).
Humid volume is the actual volume occupied by the
air/vapor combination per pound (or kilogram) of dry air. On
the psychrometric chart, HV lines represent this value (see
Figure 3-11). Humid volume is used to determine changes in
airflow rate within a ventilation system due to combining
gases of different properties, or when evaporative cooling
occurs within the system.
 = (ṁH2O)/(ṁda) = (mH2O)/(mda)
[3.31a]
It is most important to understand that the reciprocal of
humid volume is not density. The actual density of the moisture-laden air can be calculated by knowing the HV and
humidity ratio ():
ρact = (1 + ω)/HV
[3.33]
where:
ṁH2O = mass flow rate of water vapor, lbm/min [kg/s]
ρact = density of the moisture-laden airstream,
lbmmix/ft3mix [kgmix/m3mix]
ṁda = mass flow rate of dry air, lbm/min [kg/s]
mH2O = mass of water vapor, lbm [kg]
 = humidity ratio, lbmH2O/lbmda [kgH2O/kgda]
mda = mass of dry air, lbm [kg]
HV = humid volume, ft3mix/lbmda [m3mix/kgda]
Using Equation 3.4, Equation 3.31a can be written as:
 = [(nH2O)/(nda)][(MWH2O)/(MWda)]
where:
[3.31b]
where:
nH2O = number of moles of water vapor,
lbmol [gmol]
EXAMPLE PROBLEM 3-3 (Psychrometric Chart) (IP Units
Only)
MWH2O = molecular weight of water vapor,
lbm/lbmol [g/gmol]
One pound of dry air has a dry bulb temperature (Tdb) of 100
F. Moisture is added to the dry air and the wet bulb temperature
(Twb) of the moisture-laden air is 77 F. Use a psychrometric
chart to determine the following:
nda = number of moles of dry air, lbmol [gmol]
MWda = molecular weight of dry air, lbm/lbmol
[g/gmol]
•
Humidity ratio, ω
•
Dew point temperature, Tdp
ENTHALPY (h)
•
Humid volume, HV
The enthalpy (or total heat) of an ideal gas represents the
sum of the specific enthalpies for dry air (i.e., sensible heat)
and moisture (i.e., latent heat). The enthalpy of an airstream is
calculated using:
•
Density factor, df
•
Enthalpy, h
h = 0.240(T) + [1,061 + (0.444)(T)]
[3.32] IP
[h = 1.006(T) + [2501 + (1.860)(T)]
[3.32] SI
where:
h = enthalpy, BTU/lbmda [J/kgda]
 = humidity ratio, lbmH2O/lbmda [kgH2O/kgda]
T = dry bulb temperature, F [C]
Enthalpy is important when combining airstreams with dif-
Solution:
Humidity ratio (ω) is determined by first identifying the operating point defined by the intersection of the dry bulb and wet
bulb temperature lines. From this point, draw a horizontal line
to the vertical humidity ratio axis on the right side of the psychrometric chart (see Figure 3-12):
Operating Point = intersection of Tdb = 100 F
and Twb = 77 F
ω = 0.016 lbmH2O/lbmda
Dew point temperature (Tdp) is determined by drawing a horizontal line from the operating point to its intersection with the
saturation curve on the left side of the psychrometric chart:
3-16
Industrial Ventilation
FIGURE 3-12. The psychrometric chart for Example Problem 3-3
Tdp = 69 F
Humid volume (HV) is determined by identifying the humid
volume line on the psychrometric chart that intersects with the
operating point:
HV ≈ 14.45 ft3mix/lbmda
Actual density (ρact) of the moisture-laden air can be calculated by using Equation 3.33:
ρact
= (1 + ω)/HV
= (1 + 0.016)/14.45
= 0.070 lbmmix/ft3mix
Density factor is determined by using Equation 3.10:
df = (ρact)/(ρstd) = (0.070)/(0.075) = 0.93
The enthalpy (h) of the moisture-laden air can be read from
the psychrometric chart by determining the total heat line that
intersects the operating point:
EXAMPLE PROBLEM 3-4 (Moisture Level by Weight)
Moisture level is sometimes expressed in terms of volume
instead of weight. In such cases, one must be able to convert
from volume-based units to weight-based units. Assume an airwater mixture is 15% moisture water by volume. Determine the
moisture level by weight (lbmH2O/lbmda) [kgH2O/kgda].
Solution:
From ideal gas law (Equation 3.3): PV = nRuT
For air in the combination: PVda = ndaRuT
For water in the combination: PVH2O = nH2ORuT
The partial volumes to yield the mixture volume (Vmix), and T
and P are the same for both air and water. Therefore:
PVmix = (nda + nH2O)RuT
Using Equation 3.4 (n = m/MW):
h = 41.7 BTU/lbmda
mair = (nair)(MWair)
Enthalpy can also be calculated by use of Equation 3.32 IP:
mair = (nH2O)(MWH2O)
h = 0.240(T) + ω [1,061 + (0.444)(T)]
= 0.240(100) + (0.016)[1,061 + (0.444)(100)]
MWda = 28.9 lbm/lbmol [g/gmol]
MWH2O = 18.0 lbm/lbmol [g/gmol]
= 41.7 BTU/lbmda
The answer is independent of the temperature or pressure
of the combination.
Principles of Airflow
3.9.2 Temperature and Humidity Control. Temperature
and/or humidity control are frequently encountered in heating,
ventilation, and air conditioning (HVAC) applications. In
some cases, it is desirable to perform more than one psychrometric process at the same time such as: 1) cooling and dehumidification or 2) heating and dehumidifying. These psychrometric processes are depicted on the psychrometric chart in
Figure 3-13.
Use of the conservation of mass and energy equations are
pertinent to the HVAC processes presented above.
Psychrometric charts are useful for determining physical properties at each state or to determine the graphical solution to an
application.
Psychrometric processes frequently encountered when evaluating industrial ventilation ductwork problems are presented
below. These include: 1) combining airstreams at different
conditions, and 2) cooling and humidifying air. Other temperature and humidity control processes, while important to
HVAC applications, are used infrequently when designing
ductwork systems. Those seeking more information on these
other psychrometric processes are referred to Chapter 1 of
ASHRAE’s Fundamentals handbook.(3.2)
3.9.2.1 Combining Airstreams at Different Conditions. The
combining of two quantities of air at different temperatures
and moisture contents frequently occurs in duct systems.
3-17
Another application involves treating hot process gases with
outdoor air to cool the gas stream down to the desired temperature condition at the inlet of an air pollution control device.
These combining processes are assumed to occur at adiabatic
conditions (i.e., no heat transfer). Figure 3-14 depicts the process
of combining two airstreams on the psychrometric chart.
As per Section 3.7, the mass flow rate of both air and water
vapor must be conserved:
ṁ1 + ṁ2 = ṁ3
[3.34]
Remembering that ω = (ṁH2O)/( ṁda):
(ṁda-1)(ω1) + (ṁda-2)(ω2) = (ṁda-3)(ω3)
[3.35]
Energy must also be conserved:
(ṁ1)(h1) + (ṁ2)(h2) = (ṁ3)(h3)
[3.36]
When no moisture is present in the airstream, Equation 3.36
can be written as:
(ṁda-1)(Cp)(T1) + (ṁda-2)(Cp)(T2) = (ṁda-3)(Cp)(T3)
[3.37]
where:
ṁda = mass flow rate of dry gas, lbmda/min [kgda/min]
Cp = specific heat of dry gas, BTU/lbm-R [kJ/kg-K]
T = dry bulb temperature, R [K]
FIGURE 3-13. Temperature and humidity control processes plotted on a psychrometric chart
3-18
Industrial Ventilation
FIGURE 3-14. Psychrometric process of combining of two airstreams
Note that when using Equation 3.37, the units for temperature
must be in absolute units (i.e., Rankine in IP units or Kelvin in
SI units). The temperature for a gas stream consisting of air
may be in units of Fahrenheit or Kelvin.
In a process involving the combining of gas streams composed of air, Cp cancels out of Equation 3.37:
(ṁda-1)(T1) + (ṁda-2)(T2) = (ṁda-3)(T3)
[3.38]
Solution:
The density factor of both airstreams must be determined (see
Section 3.5).
Density Factor – Airstream 1
Elevation: dfe = [1  ((6.73)(10-6)(3,000))]5.258 = 0.90
Temperature: dft = (460 + 70)/(460 + 400) = 0.62
Moisture: dfm = (1 + 0.20)/(1 + 1.607(0.20)) = 0.91
df1 = (dfe)(dft)(dfm) = (0.90)(0.62)(0.91) = 0.51
Density Factor – Airstream 2
EXAMPLE PROBLEM 3-5 (Combining Airstreams at
Different Conditions) (IP Units Only)
A hot, moist gas airstream (Airstream 1) and outside, winter air
(Airstream 2) are combined to form Airstream 3. Airstream 1 is
flowing at 19,000 acfm and has a dry-bulb temperature of 400
F; it contains 0.20 pounds of water per pound of dry air.
Airstream 2 is flowing at 11,000 acfm, has a temperature of 20
F, and contains virtually no moisture. The plant is located at an
elevation of 3,000 feet ASL. Determine the final conditions of
the mixed Airstream 3.
Elevation: dfe = [1  ((6.73)(10-6)(3,000))]5.258 = 0.90
Temperature: dft = (460 + 70)/(460 + 20) = 1.10
df2 = (dfe)(dft) = (0.90)(1.10) = 1.00
The standard flow rate must now be determined for both
airstreams (see Equation 3.22b):
Airstream 1: Qstd-1 = (19,000)(0.51)/(1 + 0.20) = 8,075 scfm
Airstream 2: Qstd-2 = (11,000)(1.00)/(1 + 0) = 11,000 scfm
Now determine the mass flow rate for each airstream’s compo-
Principles of Airflow
nents (see Equation 3.20):
Airstream 1 – dry air: ṁda-1
= 605.6 lbmda/min
= (0.075 lbm/ft3)(8,075 ft3/min)
Airstream 1 – water: ṁH2O-1 = (0.20 lbmH2O/lbmda)
(605.6 lbmda/min) = 121.1 lbmH2O/min
Airstream 2 – dry air: ṁda-2 = (0.075 lbm/ft3)
(11,000 ft3/min) = 825 lbmda/min
Airstream 2 – water: ṁH2O-2 = 0 lbmH2O/min
Using the conservation of mass, the Airstream 3 mass flow
rates for dry air and water vapor are:
ṁda-3 = ṁda-1 + ṁda-2 = 605.6 + 825 = 1,430.6 lbmda/min
ṁH2O-3 = ṁH2O-1 + ṁH2O-2 = 121.1 + 0 = 121.1 lbmH2O/min
The humidity ratio of Airstream 3 is:
3 = (121.1)/(1,430.6) = 0.085 lbmH2O/lbmda
The standard airflow rate for Airstream 3 is:
Qstd-3 = (ṁda-3)/(rstd) = (1,430.6)/(0.075) = 19,075 scfm
The energy for Airstream 3 is determined using only the dry air
mass flow rates for Airstreams 1 and 2 (see Equation 3.36):
(605.6)(h1) + (825)(h2) = (1,430.6)(h3)
The enthalpy for Airstreams 1 and 2 is determined by identifying the enthalpy value on the psychrometric chart that intersects the operating point defined by each airstream’s dry bulb
temperature and humidity ratio:
h1 ≈ 340 BTU/lbmda
h2 ≈ 4.8 BTU/lbmda
Now the enthalpy for Airstream 3 can be determined:
h3 = [(605.6)(340) + (825)(4.8)]/(1,430.6) =
146.7 BTU/lbmda
Knowing the enthalpy and humidity ratio of Airstream 3, its dry
bulb temperature (T3) can be determined using Equation
3.32IP:
T3 = [146.7 + (0.085)(1061)]/[0.240 +
(0.085)(0.444)] = 203 F
The density factor for elevation and temperature for Airstream
3 can now be determined:
3-19
3.9.2.2 Cooling and Humidifying Air. The addition of water
to an airstream, such as in a wet scrubber or evaporative air
cooler, both cools and humidifies the exiting airstream. This
process can be useful for cooling hot airstreams and is known
as an adiabatic or evaporative cooling process as there is no
gain or loss of heat to the surroundings. The energy for evaporating the cold water comes from the hot airstream; the evaporation of the cold water draws heat away and cools the hot
airstream. As such, the cooling and humidification of the air
occurs at constant enthalpy. In evaporative cooling, the wet
bulb temperature of the cooling water equals the wet bulb temperature of the airstream.
The evaporative cooling process is shown in Figure 3-15 on
a sample psychrometric chart. The wet bulb temperature on
the saturation curve represents the minimum temperature to
which the air can be cooled. The humidifying efficiency represents the degree to which the exiting stream’s dry bulb temperature approaches the minimum wet bulb temperature. If a
collector takes an airstream to complete adiabatic saturation, it
is said to have a humidifying efficiency of 100%.
The humidifying efficiency of a given device may be
expressed by either of the following equations:
[3.39a]
where:
ηn = humidifying efficiency, %
T1 = dry-bulb temperature at collector inlet, F [C]
T2 = dry-bulb temperature at collector outlet, F [C]
Ts = adiabatic saturation temperature, F [C]
or:
[3.39b]
where:
ω1 = moisture content at inlet, lbmH2O/lbmda [kgH2O/kgda]
ω2 = moisture content at outlet, lbmH2O/lbmda [kgH2O/kgda]
ωs = moisture content at adiabatic saturation conditions,
lbmH2O/lbmda [kgH2O/kgda]
Density Factor – Airstream 3
Elevation: dfe = [1  ((6.73)(10-6)(3,000))]5.258 = 0.90
Temperature: dft = (460 + 70)/(460 + 203) = 0.80
Moisture: dfm = (1 + 0.085)/(1 + 1.607(0.085)) = 0.95
df3 = (dfe)(dft)(dfm) = (0.90)(0.80)(0.95) = 0.68
Now, the actual flow rate of Airstream 3 can be determined
using Equation 3.22a:
Qact = 19,075(1 + 0.085)/(0.68) = 30,440 acfm
EXAMPLE PROBLEM 3-6 (Evaporative Cooling in Wet
Scrubber) (IP Units Only)
Airstream 1 moves 10,000 acfm at a dry-bulb temperature of
200 F. This airstream possesses a moisture content of 5% and
enters a wet scrubber at an SP of -6 "wg. The wet scrubber
sprays the airstream with water and it exits as Airstream 2 with
a humidifying efficiency of 95%. The pressure drop across the
wet scrubber is 6 "wg. The plant is located at an elevation near
sea level. Determine the final conditions of Airstream 2.
3-20
Industrial Ventilation
FIGURE 3-15. Psychrometric process of evaporative cooling
Solution:
Ts = 108 F
Determine the humidity ratio of Airstream 1 by using Equation
3.31b:
s = 0.056 lbmH2O/lbmda
 = [(0.05)/(0.95)][(18.01)/(28.948)] = 0.033 lbmH2O/lbmda
Determine the density factor of Airstream 1:
dfp = [407 + (-6)]/(407) = 0.99
The saturation humidity ratio of Airstream 2 (s) can also be
determined by use of Equation 3.39b:
2 = 1 + (hn/100)( s  1) = 0.033 
(95/100)(0.056  0.033) = 0.055 lbmH2O/lbmda
Determine Qstd for Airstream 1 using Equation 3.22b:
The dry bulb temperature of the humidified Airstream 2 (T2) can
be determined by drawing a constant temperature line from the
operating point defined by the intersection of the 95% relative
humidity line and the Ts = 108 F line. Equation 3.39a can also
be used to determine T2:
Qstd = ((10,000)(0.78)/1 + 0) = 7,800 scfm
T2 = T1  (hn/100)(T1  Ts) = 200  (95/100)(200  108)
dft
= (460 + 70)/(460 + 200) = 0.80
dfm = (1 + 0.033)/(1 + 1.607(0.033)) = 0.98
df1
= (dfp)(dft)(dfm) = (0.99)(0.80)(0.98) = 0.78
Using Equation 3.20:
ṁda = (Qstd)(rstd) = (7800)(0.075) = 585 lbda/min
Using a psychrometric chart, Ts (or Twb) can be determined
by drawing a line through the operating point for Airstream 1
defined by T1 = 200 F and 1 = 0.033 lbmH2O/lbmda to the saturation curve at constant Ts. The saturation humidity ratio (s)
is determined by drawing a line from the Ts on the saturation
curve at constant  to the  scale.
= 113 F
The density factor of Airstream 2 is now determined:
dfp = [407 + (-12)]/(407) = 0.97
dft = (460 + 70)/(460 + 113) = 0.92
dfm = (1 + 0.055)/(1 + 1.607(0.055)) = 0.97
df2 = (dfp)(dft)(dfm) = (0.97)(0.92)(0.97) = 0.87
Given that the df2 already accounts for moisture in Airstream 2,
Principles of Airflow
Q is determined by use of Equation 3.20:
Qact = (7800)(1 + 0.056)/(0.87) = 9,468 acfm
The mass of water added to Airstream 2 to achieve a 95%
humidification efficiency can now be determined:
ṁH2O = (ṁda)( 2  1) = (585)(0.056  0.033)
= 13.5 lbmH2O/min
3.10
DEW POINTS
Many industrial processes release significant amounts of
water into the airstream. When a moist airstream cools as it
flows through a ventilation system, it can cool to the dew point
temperature. This cooling causes moisture in the air to condense into liquid water droplets.
The combustion of organic fuels with atmospheric air generates flue gases composed of gaseous carbon dioxide (CO2),
water vapor (H2O), nitrogen (N2), and excess oxygen (O2)
remaining from the intake combustion air. Flue gases also contain gaseous air pollutants consisting of sulfur oxides (SO2 and
SO3) and nitrogen oxides (NO and NO2). These sulfur and
nitrogen oxides form acid gases and, when cooled below the
acid dew point temperature, the flue gas can become saturated
with gaseous acid resulting in liquid acid drops forming in the
airstream.
3-21
When designing a ventilation system, the designer must be
able to understand how moisture and acid gases affect the dew
point temperature of the airstream. It is very important to manage the temperature of the airstream so that it does not lead to
the formation of liquid water or acid in the air pollution control
system.
The accumulation of liquid water droplets in ductwork may
absorb or bind with the contaminant and form a sludge on the
inside surfaces of the ductwork. This type of buildup can also
accumulate and blind the baghouse filters. Moisture problems
also cause higher static pressure losses in the ductwork and
across the filters; this leads to maintenance problems and
unacceptable system airflow performance.
At the acid dew point temperature of the flue gas, liquid acid
droplets severely attack carbon steel materials causing corrosion problems. When the temperature of an exposed metal surface is below the dew point temperature of the acid gas, serious corrosion problems may occur on uninsulated carbon steel
surfaces.
REFERENCES
3.1
Boyers, A.: Private Communication to G. Lanham
(April 2005).
3.2
ASHRAE: Fundamentals, Chapter 1, Psychrometrics
(2013).
Chapter 4
INDUSTRIAL VENTILATION SYSTEM DESIGN PRINCIPLES
NOTE: Equations with notation followed by (IP) are designated for inch-pound system only; equations followed by (SI) are designated for
metric use only. If equation bears neither, then it applies to both systems.
4.1
4.2
4.3
4.4
4.5
ADMINISTRATION OF INDUSTRIAL
VENTILATION SYSTEM DESIGN . . . . . . . . . . . . . . .4-2
DRAWINGS AND SPECIFICATIONS . . . . . . . . . . . . .4-2
DESIGN OPTIONS FOR INDUSTRIAL
VENTILATION SYSTEMS . . . . . . . . . . . . . . . . . . . . . .4-3
4.3.1 Dilution Versus Local Exhaust Ventilation
Design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4-3
4.3.2 Discharge of Emissions . . . . . . . . . . . . . . . . . . .4-3
4.3.3 Local Exhaust Ventilation System
Orientation . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4-4
DESIGN PROCEDURES . . . . . . . . . . . . . . . . . . . . . . . .4-4
4.4.1 Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . .4-4
4.4.2 Preliminary Steps . . . . . . . . . . . . . . . . . . . . . . . .4-5
4.4.3 Determining the Combustibility of Dust . . . . . .4-6
4.4.4 Calculation Methods to Optimize Design . . . . .4-6
4.4.5 Design Calculations to Estimate System
Performance . . . . . . . . . . . . . . . . . . . . . . . . . . . .4-7
4.4.6 Selection of Duct Velocities . . . . . . . . . . . . . . . .4-7
DISTRIBUTION OF AIRFLOW IN DUCT
SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4-7
Figure 4-1
Figure 4-2
Figure 4-3
Figure 4-4
Figure 4-5
Figure 4-6
Organizational Flow Chart . . . . . . . . . . . . . . . . . .4-2
Drawing with Minimum Dimensions . . . . . . . . .4-3
Drawing with Detailed Dimensions . . . . . . . . . . .4-4
Dilution or General Ventilation . . . . . . . . . . . . . .4-5
Local Exhaust Ventilation System . . . . . . . . . . . .4-6
On-Line Design (Single Fan and/or
Collector for Single or Small Group
of Contaminant Sources) . . . . . . . . . . . . . . . . . . .4-7
Table 4-1
Relative Advantages and Disadvantages of
Balance-by-Design Versus Blast Gate
Methods . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4-9
4.5.1 Balance-by-Design Method . . . . . . . . . . . . . . .4-10
4.5.2 Blast Gate/Orifice Plate Method . . . . . . . . . . .4-10
4.5.3 Adjustable Local Exhaust Systems . . . . . . . . .4-10
4.6 LOCAL EXHAUST VENTILATION
SYSTEM TYPES . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4-10
4.6.1 Tapered Main Versus Plenum Design . . . . . . .4-10
4.6.2 Plenum Design Advantages and
Disadvantages . . . . . . . . . . . . . . . . . . . . . . . . . .4-11
4.6.3 Plenum System Design Considerations . . . . . .4-12
4.6.4 Tapered Main Design Considerations . . . . . . .4-12
4.7 SYSTEM REDESIGN . . . . . . . . . . . . . . . . . . . . . . . . . .4-12
4.8 SYSTEM COMPONENTS . . . . . . . . . . . . . . . . . . . . . .4-12
4.8.1 Hoods . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4-13
4.8.2 Duct Systems . . . . . . . . . . . . . . . . . . . . . . . . . .4-14
4.8.3 Air Pollution Control Devices . . . . . . . . . . . . .4-14
4.8.4 Fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4-15
4.8.5 Exhaust Stacks . . . . . . . . . . . . . . . . . . . . . . . . .4-15
4.9 LOCAL EXHAUST VENTILATION SYSTEM
TESTING AND BALANCING . . . . . . . . . . . . . . . . . .4-15
REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4-15
Figure 4-7
Figure 4-8
Figure 4-9
Single Line Isometric Sketch of Local Exhaust
Ventilation System . . . . . . . . . . . . . . . . . . . . . . . .4-8
Plenum Duct System . . . . . . . . . . . . . . . . . . . . .4-11
Types of Plenum Duct Designs . . . . . . . . . . . . .4-13
4-2
4.1
Industrial Ventilation
ADMINISTRATION OF INDUSTRIAL VENTILATION
SYSTEM DESIGN
Industrial ventilation projects often require compliance with
occupational and/or environmental regulations, as well as various building codes and local ordinances. As such, the success
of such projects relies on the proper communication of ideas,
responsibilities, and expectations, and establishment of appropriate proofs of performance. The formation and support of a
project team is often critical to the implementation of an effective industrial ventilation system.
Figure 4-1 depicts a flow chart identifying the various
groups and their communications that may be associated with
the design and installation of industrial ventilation projects.
The size and make-up of the project team is based on the
nature of the materials controlled, size and complexity of the
project, and company size; in some cases, outside technical
expertise may be necessary. The project team is responsible for
establishing/overseeing the project requirements and design
basis, as well as system installation and commissioning.
The owner must approve of the design basis prior to the
installation phase of the project. This ensures contractors and
vendors can appropriately bid the project. A thorough design
basis also aids in minimizing potentially costly change order
requests.
It is important that the project team documents its decisions
and communications. Methods and guidelines for the organization of project teams and design management, as well as tools
for developing the design basis, are discussed in Chapter 2.
4.2
DRAWINGS AND SPECIFICATIONS
Communication of the design intent is usually made via
permanent records such as drawings, specifications, and written scope documents. The level of design detail is determined
FIGURE 4-1. Organizational flow chart
during the design basis phase. Drawings may vary from basic
single line sketches to detailed computer aided design (CAD)
drawings that include isometric views and scale models.
Specifications and other instructions may simply be typed
documents included as notes on the drawings, or they may be
a file that can be read by an engineering software package.
The level and presentation of the detailed design should
consider the needs, level of sophistication, and experience of
the intended audience. Single line sketches may be suitable for
experienced fabricators and installers. Less experienced installers may need all dimensions and other details shown. The project manager and team should know and communicate the
requirements for the level of detail through the design basis.
Late changes lead to increased costs, so a thorough review by
multiple stakeholders is recommended before completing the
design basis phase.
Once the design basis is complete, the design team should
keep the direction of the project focused and avoid any
attempts to increase the scope of the project. This will assist in
controlling cost over-runs and change orders that could delay
implementation of the design.
Experienced designers may be able to use CAD techniques
or templates to reduce drawing time. Lasers and other
resources are also available to develop more detailed field
measurements. Hence, less detailed drawings may include all
the dimensions necessary for installation and may eliminate
duplication. An example of a drawing depicting minimal
dimensioning is shown in Figure 4-2. One set of dimensions
shows only the distance between pieces of equipment. It
allows some flexibility by the installer to choose duct lengths
and flange locations to suit their installation techniques but
still meet the requirements of the design.
Figure 4-3 shows every piece dimensioned in detail. Such
detail may be necessary on projects where: 1) there are specif-
Industrial Ventilation System Design Principles
4-3
FIGURE 4-2. Drawing with minimum dimensions
ic connection requirements; 2) special duct routing is required
to ensure clearance from predicted obstructions; 3) all duct
segments are fabricated off-site and the installation is on a tight
schedule; or 4) the installers have limited experience and the
design intent should be correspondingly more explicit.
Detailed dimensioning may cost extra money if companies are
held to exact dimensions as displayed. The provision of such
detail assumes that the designer knows the best and most costeffective location of all pieces and flanges. System components that usually require detailed design information include
structural supports for duct and hoods, fire suppression equipment locations, and other features required to meet codes and
regulations.
4.3
DESIGN OPTIONS FOR INDUSTRIAL
VENTILATION SYSTEMS
The following information contains recommendations and
experiences based on good engineering practice. Note that the
most restrictive code, regulation, or specification will supersede any recommendation provided herein. For information on
the operation and maintenance of industrial ventilation systems, see the O&M Manual.
4.3.1 Dilution Versus Local Exhaust Ventilation Design.
The primary purpose of an industrial ventilation system is to
maintain a safe level of airborne contaminants by diluting
them and/or removing them from the worker’s environment.
The system components and control method(s) should be
selected for the specific process, work flow, and tasks
involved. Generally, the size and type of the equipment is
based on the process and ergonomics, size of hoods and duct-
ing (if used), and desired tradeoffs associated with reliability,
operating cost, and initial cost.
There are basically two types of ventilation systems: dilution (also called general) and local exhaust. Dilution ventilation mixes large amounts of clean air with contaminated air to
keep concentrations below allowable limits (see Figure 4-4).
Normally, dilution ventilation is used to control the potential
for fire or explosive conditions or to dilute odors. Dilution
ventilation can also include the control of airborne contaminants (e.g., vapors, gases, and particulates), but is limited to
less toxic contaminants. The design criteria for dilution ventilation systems are detailed in Chapter 10.
Local exhaust ventilation systems capture contaminants at
their point of generation and remove the contaminants from
the workplace through a duct system (see Figure 4-5). In addition, local exhaust ventilation systems also create a path for
exhaust streams of materials from plant processes.
The remainder of this chapter will focus on providing an
overview of local exhaust ventilation system design considerations; detailed design calculation methods are included in
Chapter 9.
4.3.2 Discharge of Emissions. In some cases, exhaust air
with low levels of contaminants can be discharged directly to
the atmosphere outside the workplace. Considerations impacting this decision include:
a) No government regulations prohibit doing so;
b) Levels are predictable and verifiable;
c) Other nuisances, like odors, are not sent into the atmos-
4-4
Industrial Ventilation
FIGURE 4-3. Drawing with detailed dimensions
phere; and
d) The discharge of the contaminants does not cause a
neighborhood nuisance.
An appropriate air-cleaning device is frequently necessary
when the air stream possesses substances of high toxicity or
that are emitted at high discharge rates. Air discharged outside
the plant should conform to both federal and local emission
standards and not exceed any permitted release levels. Details
for the selection and design of air-cleaning devices are included in Chapter 8.
In situations where both the contaminant emission levels
and toxicity are low, it may be possible to return the cleansed
air stream back to the workplace. An explanation of when and
how air can be recirculated is included in Chapter 11.
The requirements for air-cleaning devices are normally
determined by regulations at federal, state, and/or local levels.
Before beginning the design process, a determination should
be made concerning the use of air-cleaning devices and
required efficiencies or permitted discharge limits.
4.3.3 Local Exhaust Ventilation System Orientation. The
method of connecting the hood, air-cleaning device, and fan
can vary from system to system. Figures 4-2, 4-5 and 4-6 illustrate a variety of ways to integrate hood and collector systems.
Each of the illustrated systems has its own advantages and disadvantages.
The system design style may also be limited by architectural
considerations or the limitations of the physical space where
the equipment is to be located (e.g., there may be only one possible location for the air-cleaning device). Early in the design
process, even as the project team is being chosen, an audit
should determine alternatives for the physical location of
process equipment or ventilation system components.
Alternatives may be limited by location of exhaust stacks,
electrical power sources, soil or building structural conditions,
or access for removal of collected pollutants. Additional limitations may include lease or purchase agreements that limit
noise from the site.
The design basis may include these restrictions or recommendations, but many times the actual equipment locations are
being determined as the detailed design phase proceeds.
4.4
DESIGN PROCEDURES
4.4.1 Introduction. The layout of the hoods, ducting, aircleaning device(s), and fan should be properly designed. This
process is more involved than merely connecting system components. If the system is not carefully designed in a manner
that reliably ensures that all required airflow rates will be realized, adequate contaminant control may not be achieved.
Additionally, minimum transport velocities should always be
maintained in all ducts during the operation of systems handling particulates. The designer should consider initial capital
Industrial Ventilation System Design Principles
4-5
FIGURE 4-4. Dilution or general ventilation
costs, reliability, maintenance, and equipment life in the design
and/or selection of all system components.
Duct systems require large amounts of air to convey relatively small amounts of contaminant. For that reason, they are
one of the least efficient items in the plant. Careful system
design can meet the desired goals utilizing the least amount of
power and initial cost. In addition, the designer should also
consider reliability, maintenance, and equipment life.
Calculation procedures detailed in Chapter 9 are used to
determine the appropriate system flow rate(s) and pressure(s),
as well as duct sizes and fan operating point. Chapters 7 and 9
describe how to select a fan based on these results.
4.4.2 Preliminary Steps. With almost all design efforts,
proper organization of data and information will simplify the
process. To coordinate design efforts with all personnel involved (including the equipment or process operator as well as
maintenance, health, safety, fire, and environmental personnel), the designer should have, at a minimum, the following
data available at the start of the design process:
1) The air stream properties associated with the system to
be designed. Failure to consider the elevation, temperature, pressure, and moisture content of the system can
result in a design that fails to meet the desired requirements. See Chapter 3 for a discussion of the principles
of airflow.
2) A layout of the operations, workroom, building (if necessary), etc. The available location(s) for the air-cleaning device and fan should be determined. It is also
important to identify the location of the final system
exhaust point (i.e., where the air exits the system, usually a stack or fan discharge). Air must be discharged
such that it does not re-enter the workspace, either
through openings in the building perimeter or replacement air unit intakes. Calculations for the proper location of the exhaust stack are included in Chapter 5,
Section 5.3.
3) A line sketch of the duct system layout, including plan
and elevation dimensions, fan and air-cleaning device
location, etc. Number, letter, or otherwise identify each
hood and duct segment(s) on the line sketch for convenience (see Figure 4-7).
Most systems, when handling particulates, will locate
the fan on the clean air side of the air-cleaning device.
Other considerations may force the location of the fan
before the collector. If possible, locate the fan closest to
pieces of equipment with high static pressure losses; this
will facilitate balancing and may result in lower operating costs. Locating the fan (and air-cleaning device) in
the center of the system (see Figure 4-5) may yield a
smaller system static pressure (SSP) requirement.
4) Use ductwork made from hard, smooth materials.
Round, plastic piping, and round or oval ductwork
made of rolled sheet metal are preferred. If flexible
duct lengths are required in the system, keep them as
short and straight as possible. Flexible duct is susceptible to sagging and excessive bending, and is usually
4-6
Industrial Ventilation
FIGURE 4-5. Local exhaust ventilation system
reinforced with ribbing or corrugation, all of which
increases static pressure losses; such losses usually
cannot be determined accurately. Additionally, the
pressure loss per foot for a straight, flexible duct section can be more than twice that for a hard, smoothwalled duct segment.
5) A design or sketch of the hood for each operation with
direction and elevation of outlet for its duct connection.
Hood sketches can be in isometric or plan and elevation views. Enough detail should be included to determine the anticipated opening sizes, location and size of
slots, and any other factors that will assist with determining appropriate airflow rates and hood static pressures.
6) Information about the details of the operation(s),
including: toxicity, worker access/use, physical and
chemical characteristics, required airflow rates at
hoods or enclosures, and required minimum duct transport velocities (see Chapter 5), hood entry losses, and
required capture velocities at each hood’s face. Special
attention should be given to room air turbulence (e.g.,
due to cross ducts, supply air delivery, and other disturbing air movement) and incompatibilities between
gases, vapors, or particulates (that may intermix in the
exhaust system to assure that they do not result in fire
or explosion hazards, destructive corrosion, or toxic
mixtures). If any mixture is incompatible, separate ventilation systems or appropriate air-cleaning devices
should be provided. Consult all latest appropriate standards, including ASTM E 2012-06, Standard Guide for
the Preparation of a Binary Chemical Compatibility
Chart, for the selection of materials of construction as
well as the design routing of ventilation systems.
7) Information relevant to the process being controlled,
such as temperature, moisture content, and elevation
(above sea level) should be provided for each hood and
duct branch.
8) The method and location of all replacement air distribution devices as they affect each hood’s performance.
The type and location of supply air fixtures can dramatically affect contaminant control by creating undesirable turbulence at the hood (see Chapter 11). Perforated
plenums or duct may provide better replacement air
distribution with fewer adverse effects on hood performance.
4.4.3 Determining the Combustibility of Dust. Many particulate substances transported in ventilation systems are combustible. The transport of combustible dusts by industrial ventilation systems requires additional diligence, assessment, and
control measures to protect employees from hazards caused by
the potential for explosions and fires. Reference Chapter 12,
Section 12.3 and comply with all OSHA requirements and
other reputable guidance available when designing systems
that will transport combustible dusts.
4.4.4 Calculation Methods to Optimize Design. The
design of a local exhaust ventilation system is a continuing
process; it does not end with the initial system calculations.
Calculations and evaluations may need to be repeated several
times including: 1) during the original conceptual design, 2)
Industrial Ventilation System Design Principles
4-7
FIGURE 4-6. On-line design (single fan and/or collector for single or small group of contaminant sources)
during the final drive speed specification from as-built drawings, and 3) when providing a tool for the air balancing technician. It is recommended that the designer also use an optimization step to: 1) identify ducts with high velocities that
could wear prematurely, and 2) analyze the branches with the
highest pressure drop so that changes can be made to reduce
system static pressure. For example, a small branch duct in a
large volume system may represent the highest static pressure
loss. By increasing the flow at the hood, making the duct larger
and reducing the friction losses in the duct, the overall system
pressure may decrease with a small total increase in total flow.
This can result in a lower overall system horsepower requirement.
Additionally, after the system is in use, it will lose some
effectiveness as dust coats the duct interior wall (changing
friction losses) and fan impellers, and collectors begin to show
wear and dust buildup. The designer should consider the conditions during the operating life of the system. For instance,
where volumetric flow, face velocities, or transport velocities
are selected from a range of provided values, the upper end of
the range should be considered if the system cannot be shut
down easily for routine maintenance.
4.4.5 Design Calculations to Estimate System Performance. The calculation methods are used primarily to engi-
neer the system (determine duct sizes, estimate static pressure
requirements for fan selection, etc.). However, the data can
also be used to predict a range of operations that can be used
to support field analysis of systems. Static pressures calculated
at branches using the methods in Chapter 9 can be used as a
start point to predict possible findings when troubleshooting
systems. Note that calculation sheet data are for system design
only and this will not duplicate the actual conditions. Hood
losses, actual duct losses after material coats the inside walls,
and other fabrication influences such as grinding of welds,
etc., will impact the actual results. Values published for losses
in system components are best estimates.
4.4.6 Selection of Duct Velocities. In systems that are
intended to carry particulates, a minimum duct (conveying)
velocity is necessary to ensure that the particulate will not settle in the duct. Conversely, when a system handling ‘clean air’
is installed in a quiet area, it may be necessary to keep velocities low to avoid excessive duct noise. When axial flow fans
are used to move air streams containing no particulates, velocities of 1,000 to 1,500 feet per minute (fpm) [5.08 to 7.62 m/s]
may be preferred. In a gas or vapor exhaust system installed in
a typical factory environment where none of these restrictions
apply, the velocity may be selected to yield the lowest annual
operating cost.
To determine the optimum economic velocity, the system
should first be designed at an assumed velocity and the total
initial costs of duct material, fabrication and installation estimated. Optional duct and operating costs can be determined
for other duct velocities for comparison. This optimum economic velocity will normally range from under 2,000 fpm
[10.16 m/s] to over 4,000 fpm [20.32 m/s]. In general, a velocity of 2,500 to 3,000 fpm [12.70 to 15.24 m/s] will result in an
equivalent total annual cost that approximately equals true
optimum. See Chapter 5 for more information regarding minimum duct (conveying) velocities.
4.5
DISTRIBUTION OF AIRFLOW IN DUCT SYSTEMS
A simple exhaust system consists of a hood, duct segments,
and special fittings leading to and from an exhaust fan; it may
4-8
Industrial Ventilation
Industrial Ventilation System Design Principles
also include an air pollution control device. A complex system
possesses an arrangement of several hoods and duct segments
connected to a common duct (i.e., the main), an air pollution
control device, and one or more fans. Whether designing a
simple exhaust system or one containing multiple hoods and
branches, the same design methods apply. However, in a
multi-branch system design, it is also necessary to properly
balance the static pressure for each duct segment at the junction.
Air will always take the path of least resistance. If the
designer makes no attempt to balance the static pressure in a
multi-branch system, a ‘natural balance’ will occur at each
junction. This ‘natural balance’ will result in an undesirable
modification to the flow rate of each segment’s hood, thereby
4-9
impacting the hood’s ability to successfully capture and convey the desired contaminant. Therefore, the designer should
take appropriate steps to balance static pressures at all branch
junctions. Properly doing so will ensure that the design airflow
at each hood does not fall to its minimum as listed in
Chapter(s) 6 and/or 13.
To accomplish the balancing of static pressures in multibranch systems, the designer may use a balance-by-design
approach or install blast gates or orifice plates. The objective
of both methods is the same: to balance static pressure and
obtain the desired flow rate at each hood in the system while
maintaining transport velocity in all duct sections. Table 4-1
shows some relative advantages and disadvantages of these
two balancing methods.
TABLE 4-1. Relative Advantages and Disadvantages of Balance-by-Design Versus Blast Gate Methods
4-10
Industrial Ventilation
4.5.1 Balance-by-Design Method. This procedure balances static pressures in a multi-branch duct system without
the use of blast gates or orifice plates. It is also called the static
pressure balance method. The designer calculates the static
pressure loss for each branch segment (based upon each
hood’s design data as well as included duct fittings and length)
up to their common junction point. Any pressure imbalance at
duct junctions is resolved by either increasing airflow rates or
modifying the duct sizes or components. This balancing
method may result in lower fan static pressures (FSP) and possibly lower horsepower requirements than those obtained by
use of the blast gate/orifice plate method.
The balance-by-design method assesses the ratio of the SP
of the governing branch (i.e., the branch whose SP is greater in
magnitude) to that of the branch with the lower magnitude SP.
If the ratio is greater than 1.2 (120%), the preferred method of
balancing the pressures involves the redesign of the branch
with the lower pressure loss. The redesign may include a
change of duct size, selection of different fittings, and/or modifications to the hood design. Note that if a redesign is used, an
additional balancing step of adjusting the airflow rate is usually required. Chapter 9 details the calculation method for this
procedure. This balancing method usually results in a higher
total airflow rate than that determined when using the blast
gate/orifice plate method.
The balance-by-design method should be used when the
local exhaust system handles highly toxic materials, the system must be protected against tampering with blast gate settings (which may consequently increase employee exposures
to toxic substances), or regulatory or consensus standards prohibit the use of blast gates. This method is highly recommended for systems that exhaust explosives, radioactive dusts, or
biological materials to minimize the possibility of accumulations in the system caused by a blast gate or orifice plate
obstruction.
4.5.2 Blast Gate/Orifice Plate Method. With this air distri-
bution method, the airflow rates of two joining ducts are
achieved by blast gate (also known as “cutoffs”) adjustments
after the installation of a properly designed orifice plate that
results in the desired static pressure balance at the junction. No
modifications to airflow rates or duct diameters or components
are required with this balancing method. A damper may be
used in the same fashion as a blast gate to balance the static
pressure of a duct segment whose installation is to be completed at a future date. This balancing method may result in a
lower total airflow rate and, theoretically, lower horsepower
requirement than that determined by using the balance-bydesign method.
The system data and design calculations for this balancing
method are the same as for the balance-by-design method,
except airflow rates, duct sizes, and fittings are not modified.
The blast gates are adjusted after installation to provide the
required static pressures at the design airflow rates. Note that
a change in any single blast gate setting in a system will impact
the airflow rates in all system branches. Also, readjusting the
blast gates during the system balancing process can sometimes
result in increases to the actual FSP and fan power requirements. Calculation methods for the employment of these balancing devices are included in Chapter 9. Similarly, orifice
plate opening sizes may be changed to reflect actual requirements at start-up or when system revisions are made.
However, orifice plate design usually infers a more permanent
installation because they are not adjustable.
With this method, the static pressure needed to balance the
branch will be the difference between the calculated static
pressures in the joining branches. In practice, many balancers
iteratively increase the insertion depth while balancing. This
can result in higher system static pressures and greater energy
use than that determined by using the balance-by-design
method. As such, the designer may wish to ensure the system’s
fan and motor are sized to account for any additional pressure
and energy capacity that may be required when using this balancing method. See Chapter 4 of the O&M Manual for a discussion of balancing methods and techniques to reduce the
total static pressure in a system balanced using blast gates/orifice plates.
Orifice plates are essentially fixed blast gates and have
many of the same advantages and disadvantages of each
method. The method of calculating orifice plate openings can
be found in other texts with varying results. The location of
blast gates and orifice plates are dependent on the location
within the duct system (near elbows and hoods or other disturbances). Five duct diameters of straight ductwork before and
after a disturbance are preferred to yield predictable results.
Losses due to blast gates (as a function of insertion depth) are
difficult to predict because of the different blade shapes and
clearances.
4.5.3 Adjustable Local Exhaust Systems. Some local
exhaust systems are designed on the assumption that only a
fraction of the total number of hoods will be operating at any
given time. In such cases, the airflow to the branches not used
will be shut off with dampers or blast gates. This practice may
lead to plugging in a tapered system when the minimum transport velocity is not maintained. This procedure is not recommended unless the minimum transport velocity can be assured
in all ducts during any variation of closed blast gates. It is better to design these systems with individual branch lines converging close to the fan inlet to minimize the lengths of duct
mains. Additionally, some NFPA standards prohibit intermittent use of blast gates as shut off valves.
4.6
LOCAL EXHAUST VENTILATION SYSTEM TYPES
4.6.1 Tapered Main Versus Plenum Design. There are two
general classes of duct system designs: tapered main systems
and plenum systems. The duct in a tapered main system gradually gets larger as airflows are merged together. If the system
Industrial Ventilation System Design Principles
transports particulate (dust, mist, or condensable vapors), a
tapered main system maintains the minimum transport velocity in all horizontal and vertical ducts. Figures 4-2 and 4-5
depict examples of tapered systems.
In a plenum exhaust system, minimum transport velocities
are maintained only in the branch ducts to prevent settling of
particulate matter. These ducts connect to an oversized plenum
where the design velocity fall well below the minimum transport velocity values (below 1,000 fpm [5.08 m/s]). The function of this plenum is to provide a low-pressure loss path for
airflow from the various branches to the air pollution control
device or the fan. This helps to maintain balanced exhaust in
all duct branches and often minimizes operating power. Figure
4-8 illustrates a plenum duct system.
In another plenum design, the duct diameter may be
increased so that particulate entrained in the air stream can settle out. Certain mist and coolant control systems are designed
this way to encourage settling of droplets in the duct, which
are then drained from the system.
Regardless of whether a tapered or plenum system is
employed, following proper calculation procedures will provide a workable system design.
4.6.2 Plenum Design Advantages and Disadvantages.
Tapered main systems represent the most common type of system for local exhaust ventilation system designs. However,
FIGURE 4-8. Plenum duct system
4-11
plenum systems offer some advantages when the handling of
mists or transport velocities are not an issue.(4.2) Both methods
have varying success based on the material(s) being collected.
Advantages of the plenum exhaust system include the following:
1) Branch ducts can be added, removed, or relocated at
any convenient point along the main; modifications
limited only by the total airflow and pressure available
at the fan. (NOTE: Systems may need to be rebalanced
every time a line change is made.) Some plenum systems are designed to automatically adjust to change in
the number of active exhaust points. For example, a
static pressure controller could be used to change a
variable frequency drive on the exhaust fan to maintain
a set static pressure.
2) Branch ducts can be closed off and the airflow rate in
the entire system reduced (if minimum transport velocities are maintained in the remaining branches).
3) The main duct can act as a settling chamber for large
particulate matter or liquids and refuse material that
might be undesirable in the air pollution control device
or fan. It is important to allow for removal of this collected material during the operation of the system
through drains, drag conveyors, etc.
Limitations of the plenum design include the following:
4-12
Industrial Ventilation
1) Sticky, linty materials tend to clog the main duct.
Buffing dust and lint are subject to this limitation and
the plenum design is not recommended for these materials.
2) Materials that are subject to direct or spontaneous combustion should be handled with care. Some materials,
such as wood dust or oil mist, have been handled successfully in plenum systems (NOTE: Ensure that
appropriate combustible dust provisions have been met
if using a plenum system to control wood dust.).
Explosive dusts such as magnesium, aluminum, titanium, or grain dusts cannot be handled in systems of this
type. Applicable NFPA and other codes may require
tapered main systems and maintenance of minimum
transport velocities in all ducts, depending on the materials handled.
4.6.3 Plenum System Design Considerations. Control
airflow rates, hoods, and duct sizes for all branches are calculated in the same manner as with tapered duct systems; such
calculations are addressed in Chapter 9. The branch segment
with the greatest pressure loss will govern the static pressure
required in the main duct and fan. Other branches will be
designed to operate at the governing static pressure by use of
the balance-by-design or blast gate method.
When the main plenum is relatively short, or the air pollution control devices or fans can be spaced evenly along the
duct, static pressure losses due to airflow in the main plenum
can be minimized. For long plenums, it may be necessary to
calculate the friction through the plenum using methods presented in Chapter 9. Design plenum velocities are usually less
than 50% of the branch velocity; they can be as low as 1,000
fpm [5.08 m/s]. Note that lower plenum velocities will result
in larger sized plenums and possibly higher initial installation
costs. Connections to air pollution control devices, fans, and
exhaust stacks are calculated in the usual manner, with consideration for maintaining minimum transport velocities.
Various types of plenum exhaust systems are used in industry (see Figure 4-9). They include both self-cleaning and manual-cleaning designs. Self-cleaning types include pear-shaped
designs that incorporate a drag chain conveyor in the bottom
of the duct. This is used to convey the particulate to a chute,
tote box, hopper, or other enclosure for disposal. Another selfcleaning design uses a rectangular main with a belt conveyor.
In both types, the conveyor may be run continuously or on
periodic cycles to empty the main duct before considerable
buildup occurs. A third type of self-cleaning design utilizes a
settled conveying main duct system to remove material; this
system usually operates continuously to avoid clogging.
Manual-cleaning designs may be built into the floor or large
enclosures behind the equipment to be ventilated. Experience
indicates that these should be generously oversized, particularly the under-floor designs, to permit added future exhaust
capacity as well as convenient access for cleanout.
4.6.4 Tapered Main Design Considerations. The tapered
main system is the standard design method used for most local
exhaust ventilation systems. In most cases, a properly
designed and sized tapered main system can provide relatively
constant velocities throughout the duct network. If these velocities meet the minimum transport velocity requirements (see
Chapter 5), particulate will be transported to the collection
device. However, the flow of any gas stream through a duct
system can result in eddies and places of high turbulence, particularly at elbows and branch entry junctions of two ducts.
Higher minimum duct velocities may be specified where
dropout of material is especially dangerous (e.g., flammable
and toxic materials). This is especially the case for extremely
long runs of duct or sections where there are several fittings
located closely together.
Less energy (horsepower) is required to operate more
streamlined systems (i.e., those using long radius elbows,
small angled branch entries, efficiently designed hoods, etc.).
While this results in a higher initial price, the cost of operating
horsepower is realized through the life of the system (sometimes 20 years or more). The designer should be cautioned to
the effects of using cheaper and less energy-efficient parts in
the system design.
4.7
SYSTEM REDESIGN
Many industrial ventilation systems undergo some form of
modification (e.g., process change, equipment relocation or
replacement, addition or deletion of a branch segment, etc.)
after having been installed. If changes are to be made to an
existing system, they should only be done as part of a management of change program. Failure to properly assess the impact
of changes to an existing system may result in poor hood performance and jeopardize employee safety and health. The
same techniques and calculation methods that are used to
design the original system are also utilized when a system is
modified. Chapter 8 in the O&M Manual, Modifying
Industrial Ventilation Systems, provides guidelines to aid with
successfully modifying industrial ventilation systems without
negatively impacting the system’s performance.
4.8
SYSTEM COMPONENTS
Once the basic system layout has been determined (see
Section 4.3.3), design of the individual system components
can be determined. Most local exhaust ventilation systems
have five components: 1) hood(s), 2) duct system, 3) air pollution control device(s), 4) air moving device(s), and 5) an
exhaust stack. Details associated with the design and specification of these components are included in Chapters 5 through
9 of this Manual. Also, many systems require the installation
of a successfully designed supply air system and/or recirculation of a cleansed exhaust air stream back into the facility; supply air systems and recirculation of exhaust air streams are
both discussed in Chapter 11.
Industrial Ventilation System Design Principles
4-13
FIGURE 4-9. Types of plenum duct designs
4.8.1 Hoods. A local exhaust hood collects contaminant
generated by a process or operation in an air or other gas
stream. These contaminants may be particulate (solid and/or
liquid) or gaseous in nature. The type of hood to be used will
depend on the physical characteristics of the process equipment, contaminant generation mechanism, and operator/
equipment interface. Hoods have a wide range of physical
configurations but are commonly grouped into three general
categories: enclosing, capturing, or receiving hoods. Chapter 6
provides a more complete discussion of hood types and design
considerations; it also contains calculation methods used to
determine the hood airflow and static resistance when other
specific guidance (such as that found in Chapter 13) is not
readily available.
4-14
Industrial Ventilation
Chapter 13 contains many hood design specification plates
grouped by major operation types. These ventilation specifications (VS) are specifically identified by a VS-plate number.
For example, VS-80-11, located in Chapter 13, provides hood
design specifications and airflow rate (Q), hood loss factor
(Fh), and minimum transport velocity (Vt) data for a grinding
wheel hood possessing surface speeds (sfpm) [sm/s] below
6,500 sfpm [32.50 sm/s]. Data from the VS-plate are used to:
specify hood dimensions; identify the airflow rate necessary
for the hood to successfully capture the contaminant; determine the energy necessary to move the contaminated air
through the hood and into the duct system; and properly size
the duct system to ensure that the minimum transport velocity
is maintained.
4.8.2 Duct Systems. After the hood design and locations
have been determined, they are connected through a duct system to the air pollution control device(s) and/or fan. Chapter 5
contains information regarding duct design principles and considerations. The method of sizing duct systems is described in
detail in Chapter 9. The Sheet Metal Contractors Association
of North America (SMACNA) provides construction standards for round and rectangular industrial ducts.(4.3, 4.4)
4.8.3 Air Pollution Control Devices. Often dusts, fumes,
and toxic or corrosive gases should not be discharged directly
to the atmosphere. To meet most regulations for air emissions,
use of an air pollution control device will be required to separate (or render harmless) the contaminants from the air stream.
The emissions to be controlled may take the form of a gas, liquid, or solid, or any combination of the three forms. Organic
and combustible vapors and water and acid mists require special considerations. Exhaust systems handling such materials
should be provided with an adequate air pollution control
device as outlined in Chapter 8 of this Manual and Chapter 6
of the O&M Manual.
Before an air pollution control device is selected, it is important to know the physical characteristics of the air stream as
well as the desired maintenance and access requirements. The
nature of the material(s) being collected, the required efficiencies, and the temperatures of the air (or gas) stream must be
known to determine the required collection method(s).
Chapter 8 discusses many available air pollution control technologies in detail. The air-cleaning device should be designed
with reliable operating parameters. Additionally, many installations require emissions monitoring or proof of continual
operation by measuring direct or surrogate conditions in the
system. This has replaced the emphasis from proof of performance at start-up with more conservative selections.
Maintenance and operating costs should also be considered
when selecting an air pollution control device. Generally, the
system can be operated through many cycles of start-up and
shut down. The air-cleaner must operate stably through these
cycles. It should be accessible for maintenance and one should
also consider if operation will be required even if there are
problems with the device. The latter would require a design
with “off-line” access so maintenance or repairs can be performed while the unit is operating.
Other important considerations include the physical size of
the equipment, installation location (inside or outside of the
plant), and the method(s) of removing and disposing of the
collected contaminants. Ultimately, the device should perform
reliably and provide the efficiencies required to meet local,
state, and federal regulations. These requirements are normally
listed in the design basis and commissioning documents. This
may include requirements for outlet loading (which is the preferred specification) or an overall efficiency rating for the unit
itself. Before any information can be included in the design
basis, careful research should be done to determine the correct
application for the emission control device and the guarantees
needed from vendors to result in a successful installation.
These contractual guarantees may also extend past the initial
installation and include maintenance and replacement parts
(e.g., filter bags, etc.) for a specified length of time.
For example, one vendor may select an air-cleaning device
that is smaller and will meet all requirements at start-up. But
operation over the life of the unit may result in a higher pressure drop and horsepower, or may require more changes of
bags or other maintenance to keep operating at required efficiencies. Life cycle costs of the selected device should assess
both electric power costs as well as ongoing operating costs.
Focusing on initial cost only may result in a financial burden
borne for the remaining life of the system. See Chapter 2 for
information on system cost considerations.
The designer must consider the change in pressure drop
(over time) of the collection device in many cases. If a system
is started with clean bags and is not seeded with a pre-coat,
then filter pressure drop across the bag media (ΔP), expressed
in "wg [Pa], may be extremely low and initial airflows at the
hoods may be higher than desired. This can have a negative
effect on the operation of the system because the higher velocities through the media can embed particles in spaces between
the media fibers and retard effective cleaning. The system may
also be connected to a process where high airflows have a negative impact. Similarly, a high initial airflow may give false
airflow readings as the system is started and balanced.
Pre-coating of bags may be the best solution to reduce the
impact of high fluctuations in pressure drop. Alternatively,
artificial resistance could be added to the fan by employing an
outlet damper possessing a feedback circuit to provide a constant inlet static pressure to the dust collector. The use of a
variable frequency drive (VFD) is another possible solution;
prices for these devices have dropped since their development,
making them an affordable option. (Note: If a VFD or inlet fan
damper is used for volume control, remember that minimum
transport velocities must be maintained in the duct system.)
Regardless of the air pollution control device selected, the
design should always be able to provide the desired airflow at
Industrial Ventilation System Design Principles
the maximum pressure drop encountered (i.e., baghouse at
maximum ΔP).
4.8.4 Fans. To move air in a local exhaust ventilation system, energy is required to overcome the system losses. These
system losses are caused by dynamic and frictional pressure
losses associated with moving air into and/or through the
hood(s), duct components, air pollution control device(s), and
any system effect losses associated with poor fan inlet and/or
outlet conditions (see Chapter 7, Section 7.4). A properly
selected, installed, and maintained fan functions as a differential pressure generator; it supplies the energy to overcome the
system pressure losses and provides the desired airflow rate at
the hood(s).
Fans (also called blowers) are the primary air moving
devices used in industrial ventilation applications. A less frequently used air moving device is the ejector. An ejector is
used when it is not desirable to pass air containing corrosive,
flammable, explosive, hot, sticky, or other troublesome materials directly through a fan. Ejectors are extremely inefficient
and generally possess higher noise levels than fans.
Fans can be divided into three basic design types: axial, centrifugal, and specialty. Generally, axial fans are used to provide
airflow at lower resistances while centrifugal fans are used to
provide airflow at higher resistances. In most cases, axial fans
are used for clean air applications; although, there are special
axial fan designs that can handle air streams with minimal
amounts of particulate. Centrifugal fans are often used in
many industrial ventilation systems due to their ability to
move air containing significant particulate loading or objects
such as cans, wood chips, etc.
Selection of an appropriate air moving device can be a complex task. The designer is encouraged to take advantage of all
available information from both applicable trade associations
and individual manufacturers. Chapter 7 discusses the characteristics and design considerations for the selection of the correct type of air moving device for a given local exhaust ventilation system. The Air Movement and Control Association
(AMCA) certifies fan performance and also provides numerous publications regarding fans, including their specification
and performance.(4.5)
4.8.5 Exhaust Stacks. The final component of an industrial
ventilation system is the exhaust stack; an extension of the
exhaust duct above the roof or grade. Assuming all exhaust
emission levels are met and maintained, two design considerations impact the placement of an exhaust stack for a local
exhaust ventilation system. First, the exhausted air stream
should escape the building envelope so that it does not return
directly through replacement and/or heating, ventilation, and
air conditioning (HVAC) systems. Then, once the air stream
has escaped the building envelope, the stack should provide
sufficient dispersion so that the plume does not cause an unacceptable situation when it reaches the ground. See Chapter 5
for a further discussion of exhaust stack design.
4-15
In some situations, the cleansed air stream can be recirculated to the plant. See Chapter 11 for guidance on recirculated air.
4.9
LOCAL EXHAUST VENTILATION SYSTEM
TESTING AND BALANCING
The exhaust system should be tested and balanced before
operation (see Chapter 3 of the O&M Manual). Openings for
sampling should also be provided in the discharge stack and/or
duct network to test for compliance with air pollution codes or
ordinances. Test ports should be located as required to verify
both fan and duct system flow and pressure. Additionally, if
employees are to work in close proximity to local exhaust
hoods, personal air monitoring should be performed to ensure
that the exhaust system achieves its primary goal of protecting
their safety and health.
REFERENCES
4.1
Hemeon, W.L.C.: Plant and Process Ventilation, 3rd
Edition, pp. 215–218. Lewis Publishers (1999).
4.2
Air Force: AFOSH Standard 161.2 (1977).
4.3
Sheet Metal and Air Conditioning Contractors’
National Assoc., Inc.: Round Industrial Duct
Construction Standards. Tysons Corner, Chantilly, VA
(2017).
4.4
Sheet Metal and Air Conditioning Contractors’
National Assoc., Inc.: Rectangular Industrial Duct
Construction Standards. Tysons Corner, Chantilly, VA
(2011).
4.5
Air Movement and Control Association, Inc.: AMCA
Standard 210-16, Laboratory Methods of Testing Fans
for Certified Aerodynamic Performance Rating.
Arlington Heights, IL (2016).
Chapter 5
DUCT SYSTEM AND DISCHARGE STACK DESIGN
PRINCIPLES
NOTE: Equations with notation followed by (IP) are designated for inch-pound system only; equations followed by (SI) are designated for
metric use only. If equation bears neither, then it applies to both systems.
5.1
5.2
DUCT SYSTEMS AND DISCHARGE STACKS . . . . .5-2
DUCT CONSTRUCTION CONSIDERATIONS . . . . . .5-2
5.2.1 Duct Sizing and Minimum Transport Velocity .5-2
5.2.2 Materials of Construction . . . . . . . . . . . . . . . . .5-3
5.2.3 Duct Fabrication Methods . . . . . . . . . . . . . . . . .5-3
5.2.4 Fabrication Standards for Materials
Other Than Steel (IP Units) . . . . . . . . . . . . . . . .5-5
5.2.5 Duct Component Considerations . . . . . . . . . . . .5-5
5.2.6 Ancillary Equipment Design Considerations . .5-6
5.3 DISCHARGE STACKS . . . . . . . . . . . . . . . . . . . . . . . . .5-6
REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .5-10
____________________________________________________________
Figure 5-1
Figure 5-2
Figure 5-3
Figure 5-4
Figure 5-5
Figure 5-6
Effects of Building on Stack Discharge . . . . . . . .5-7
Effective Stack Height . . . . . . . . . . . . . . . . . . . . . .5-8
Wake Downwash Effects . . . . . . . . . . . . . . . . . . . .5-9
Stackhead Design . . . . . . . . . . . . . . . . . . . . . . . . .5-11
Rain Caps . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .5-11
Principles of Duct Design Elbows . . . . . . . . . . . .5-12
Figure 5-7 Heavy Duty Elbows . . . . . . . . . . . . . . . . . . . . . . .5-13
Figure 5-8 Cleanout Openings . . . . . . . . . . . . . . . . . . . . . . . .5-14
Figure 5-9 Principles of Duct Design . . . . . . . . . . . . . . . . . .5-15
Figure 5-10 Principles of Duct Design – Branch Entry . . . . .5-16
Figure 5-11 Principles of Duct Design – Fan Inlets . . . . . . . .5-17
Figure 5-12 Blast Gates and Cutoffs . . . . . . . . . . . . . . . . . . . .5-18
____________________________________________________________
Table 5-1
Table 5-2
Range of Minimum Duct Design Velocities . . . . .5-2
Typical Physical and Chemical Properties of
Fabricated Plastics and Other Materials . . . . . . . .5-4
5-2
Industrial Ventilation
5.1
DUCT SYSTEMS AND DISCHARGE STACKS
Properly selected and sized hoods, air pollution control
devices, fans, and motors are critical to any successful industrial ventilation system installation. However, duct systems
and exhaust stacks play an equally important role in such systems. Failure to properly select, size, and install a duct system
may result in issues such as: settled particulate buildup in the
duct, erosion or corrosion of duct components, increased
maintenance needs, collapsed or expanded duct walls due to
system pressures, and/or collapsed duct systems due to lack of
proper structural support. Improper discharge stack installation
may result in: damage to the fan due to improper structural
support; water collecting in the system due to precipitation;
and/or re-entrainment of contaminated air back into the facility
through make-up air and HVAC units and open windows and
doorways.
This chapter is devoted to the design of industrial ventilation duct systems and exhaust stacks.
5.2
DUCT CONSTRUCTION CONSIDERATIONS
Duct systems are responsible for transporting contaminated
air from the local exhaust hood(s) to the system’s air pollution
control device(s), fan(s), and discharge stack(s) or discharge(s). These systems must be made from the appropriate
material and thickness or gauge to withstand: heat; corrosion
from gases, vapors, liquids, and solids; and erosion due to
transported particulate. Additionally, proper structural support
must also be installed to withstand duct pressures as well as
gravitational forces. The designer should address the construction details including materials and methods of construction.
Ducts are specified most often for use in the low static pressure range (i.e., -20 "wg to +20 "wg) [-4982 Pa to +4982 Pa];
but higher static pressures are occasionally encountered. The
duct can also convey air or gas at high temperatures and contaminated with abrasive particulate or corrosive aerosols and
vapors. Whether conditions are mild or severe, correct design
TABLE 5-1. Range of Minimum Duct Design Velocities
and competent installation of all system components are necessary for proper functioning of any local exhaust ventilation
system.
Exhaust system components should be constructed with
materials suitable for the conditions of service and installed in
a permanent and workmanlike manner. To minimize friction
loss and turbulence, the interior of all ducts should be smooth
and free from obstructions, especially at connections between
components. Ideally, designers should consult Sheet Metal
Contractors’ Association of North America (SMACNA) or
other industry standard sources to ensure the appropriate
design and selection of duct components.(5.1,5.2)
5.2.1 Duct Sizing and Minimum Transport Velocity.
When selecting the appropriate diameter for a duct segment,
one must consider the minimum transport (or conveying)
velocity. The minimum transport velocity represents the velocity that must be maintained in the duct system to ensure that
particulate contaminants do not settle to the bottom of the duct.
Settled particulate poses a combustible dust hazard, may completely plug the duct, and can lead to structural failure of the
duct system. Table 5-1 presents a range of minimum duct
velocities for a variety of contaminants. Note that every particulate is aerodynamically different and that a prudent designer
should be conservative in their selection of proper duct velocities. Such velocities must be able to convey both fine and
coarse particulate and ensure re-entrainment if, for any reason,
particulate settles out in any portion of the duct system prior to
reaching an air pollution control device. Note that any duct
velocity may be for design purposes when transporting a gas
or vapor in a duct system (though such systems should be evaluated for the potential for condensate formation).
When sizing the diameter of a duct segment, the designer
must first determine two things: 1) the actual airflow rate (Qact)
moving through the duct, and 2) the minimum transport velocity (Vt) necessary to keep any particulate entrained in the
airstream. Values for Qact and Vt are provided in VS-plates
Duct System and Discharge Stack Design Principles
contained in Chapter 13 for a variety of local exhaust ventilation hood designs. If such information is not available in
Chapter 13, the designer must use Table 6-3 (Chapter 6) to
determine Q and select an appropriate transport velocity from
Table 5-1.
Once the appropriate Q and Vt have been obtained, the
designer can then determine the target duct area (At) by dividing Q by Vt (i.e., At = Q/Vt). The designer then selects an available duct diameter possessing an area less than the At. If the
designer selects a duct diameter whose area exceeds the At,
then the Vt will not be maintained in the duct segment; this
may cause settling of particulate in the duct segment. Note that
the selection of non-standard duct sizes may result in increased
initial costs for duct system components; use standard sizes
when feasible to control such costs. Table 9-2 (Chapter 9) presents the areas and circumference of standard duct sizes.
5.2.2 Materials of Construction. Duct components, as well
as hoods and other fabrications, are to be constructed of black
iron or welded galvanized sheet steel (flanged and proper gaskets included), unless the presence of corrosive gases, vapors,
mists, or other conditions make such material impractical. In
those cases, stainless steel, PVC, special coatings, or some
other material compatible with the gas stream components
should be used. Arc welding of black iron lighter than 18
gauge is not recommended. Galvanized construction is not
recommended for temperatures exceeding 400 F [204 C]. It is
recommended that a specialist be consulted for the selection of
materials best suited for applications when corrosive atmospheres are anticipated. Table 5-2 provides a guide for selection
of plastic materials for corrosive conditions.
There are four classifications for exhaust systems handling
non-corrosive applications:
• Class 1 (Light Duty): for nonabrasive applications
(e.g., replacement air, general ventilation, gaseous
emissions control with no oil mist or condensing
vapors)
•
•
•
Class 2 (Medium Duty): for applications with moderately abrasive particulate in light concentrations (e.g.,
buffing and polishing, woodworking, grain dust)
Class 3 (Heavy Duty): for applications with highly
abrasive particulate in low concentrations, (e.g., abrasive cleaning operations, dryers and kilns, boiler
breeching, foundry sand handling)
Class 4 (Extra Heavy Duty): for applications with highly abrasive particles in high concentrations (e.g., materials conveying high concentrations of particulate in all
examples listed under Class 3; usually used in heavy
industrial plants such as steel mills, foundries, mines,
and smelters)
5.2.3 Duct Fabrication Methods. For most conditions,
round duct is recommended for industrial ventilation, air pollution control, and dust collecting systems. Compared to nonround duct, it provides for lower friction loss; its higher struc-
5-3
tural integrity also permits lighter gauge materials and fewer
reinforcing members. Round duct should be constructed in
accordance with SMACNA standards.(5.1) Metal thickness
(gauge) required for round industrial duct varies with classification, static pressure, reinforcement, and span between supports. Metal thicknesses required for the four classes are based
on design and use experience.
Rectangular ducts are not preferred in industrial ventilation
system applications. They should only be used when space
requirements preclude the use of round or oval duct construction. Rectangular ducts should be as nearly square as possible
to minimize resistance, and they should be constructed in
accordance with SMACNA standards.(5.2)
For limited applications, spiral wound duct is adequate and
less expensive than custom construction. However, spiral
wound duct should not be used for Classes 3 and 4 because it
does not withstand abrasion as well as smooth metal duct. It
also should not be used for applications involving the transport
of oil mists or other vapors that may condense and seep
through seams. Applications where materials may collect on
the interior surfaces, such as paper trim and stringy materials,
may also not be suitable for spiral duct.
Elbows, branch entries, and duct system components should
be fabricated, if necessary, to achieve good design (see Figures
5-6 through 5-12). Special considerations concerning the use of
spiral duct in local exhaust ventilation systems are as follows:
1) Unless flanges are used for joints, the duct should be
supported close to each joint, usually within 2 inches
[50.8 mm]. Additional supports may be needed. See
reference 5.1 or 5.2, as applicable.
2) Joints should be sealed by methods shown to be adequate for the intended service.
3) Systems can be leak tested after installation at the maximum expected static pressure. The acceptable leakage
criteria, often referred to as leakage class, should be
carefully selected based on the hazards associated with
the contaminant.
4) Fittings and elbows should be built with proper entry
angles and throat radius to duplicate round duct standards. This includes entry on the taper and not in round
duct after or before the taper.
Where condensation may occur (e.g., as in moisture-laden
air or oil mist systems, etc.), the duct system should be liquidtight. Appropriate provisions should also be made to attain
proper duct sloping and drainage. Spiral duct should not be
used for these applications.
Ducts using clamp flanges may be used for small duct operations, particularly where hoods or machines are frequently
moved, or if frequent removal for cleaning is required. This
design incorporates a quick over-center levered clamp to join
the rolled lips of all components. These duct systems can be
fabricated in stainless steel or galvanized steel and generally
are available only in small sizes (< 24" [600 mm] diameter). If
5-4
Industrial Ventilation
TABLE 5-2. Typical Physical and Chemical Properties of Fabricated Plastics and Other Materials
Duct System and Discharge Stack Design Principles
this design is used, the rolled lips for connections should be
mechanically formed on the end of the components by rolling
the duct back on itself. Such ducting is to be longitudinally
lock-seamed. Sleeves may be used for field adjustments, but
sealing of the duct should meet the standards as required for
standard SMACNA installations. There may be requirements
for more hangers to provide the same structural integrity as traditional round duct standards. Metal thickness must be at least
the same as standard round duct built to SMACNA standards.(5.1)
5.2.4 Fabrication Standards for Materials Other Than
Steel (IP Units). Equation 5.1 can be used for specifying ducts
to be constructed of metals other than steel. For a duct of infinite length, the required thickness may be determined from:
[5.1] IP
where:
t = thickness of the duct in inches
d = diameter of the duct in inches
r = intensity of the negative pressure on the duct
(lbf/in2)
E = modulus of elasticity in lbf/in2
 = Poisson’s ratio (a dimensionless material
constant)
The above equation (for Class 1 duct) incorporates a safety
coefficient that varies linearly with the diameter (d), beginning
at 4 for small ducts and increasing to 8 for duct diameters of
60 inches. This safety coefficient has been adopted by the
sheet metal industry to provide for lack of roundness, excesses
in negative pressure due to particle accumulation in the duct
and other manufacturing or assembly imperfections unaccounted for by quality control, and tolerances provided by
design specifications.(5.1)
Additional metal thickness must be considered for Classes
2, 3 and 4.
Longitudinal joints or seams should be welded. All welding
should conform to the standards established by the American
Welding Society (AWS) structural code.(5.8) Double lock seams
are limited to Class 1 applications.
5.2.5 Duct Component Considerations. Duct systems
subject to wide temperature fluctuations should be provided
with expansion joints. Flexible materials used in the construction of expansion joints should be selected with temperature
and corrosion conditions considered.
Elbows and bends should be a minimum of two gauges
heavier than straight lengths of equal diameter and have a centerline radius of at least two and preferably two and one-half
times the duct diameter (i.e., r/d = 2.0–2.5). Large centerline
radius elbows are recommended where highly abrasive dusts
are being conveyed (see Figure 5-6).
Elbows of 90° should be five-piece construction for round
5-5
duct up to six inches in diameter, and seven-piece construction
for larger diameters. Turns of less than 90° (known as “angles”)
should have a proportional number of pieces. Prefabricated
angles and elbows of smooth construction may be used.
Reinforced flat back elbows can be used where high particulate
loading is encountered (see Figure 5-7).
Where the air contaminant includes particulate that may settle in the duct, clean-out doors should be provided in horizontal runs, near elbows, junctions, and vertical runs (see Figure
5-8). The spacing of clean-out doors should not exceed 12 feet
[3.7 m] for ducts of 12 inches [305 mm] or less in diameter,
but may be greater for larger duct sizes. Removable caps
should be installed at all terminal ends and the last branch connection should not be more than six inches from the capped
end.
Transitions in mains and sub-mains should be tapered. The
taper should be at least five units long for each single unit
change in diameter (e.g., 5 inches long for every 1 inch
increase in diameter), possessing not more than a 45° included
angle (see Figure 5-9).
All branches should enter the main at the center of the transition at an angle not to exceed 45° with 30° preferred in most
cases (see Figure 5-10). Smaller angles may be specified for
abrasive materials. To minimize turbulence and possible particulate fall out, connections must be to the top or side of the
main with no two branches entering at opposite sides.
A straight duct section of at least six equivalent duct diameters should be used when connecting to a fan (see Chapter 7
for discussion of system effects). Elbows or other fittings at the
fan inlet will seriously reduce the volume discharge of the fan
(see Figure 5-11). The diameter of the inlet duct should be
approximately equal to the fan inlet diameter.
Avoid use of flexible duct especially where the formation of
severe bends is not restricted. Where required, use a non-collapsible type that is no longer than necessary to perform the
required flexibility of the connection (< 2 feet [0.6 m]). Refer
to the manufacturer’s data for friction and bend losses.
Commercially available seamless tubing for small duct
sizes (i.e., up to 8 inches) may be more economical on an
installed cost basis than other types. Plastic pipe may be the
best choice for some applications (e.g., corrosive conditions at
low temperature) but could be a bad application for abrasive
and combustible dusts.
Friction losses for duct not built to SMACNA standards can
be different than standard construction. For specific information, consult the manufacturer’s data.
Where blast gates or dampers are used, locate them at least
5 diameters away from elbows or other interferences. Ensure
that dampers cannot be adjusted after setting by locking in
place (Figure 5-12).
Hoods should be fabricated from the same materials as the
duct and a minimum of two gauges heavier than straight sections of connecting branches. They should also be free of sharp
edges or burrs and reinforced to provide necessary stiffness.
Ergonomic considerations for operator access and mainte-
5-6
Industrial Ventilation
nance should be considered in all hood designs.
Discharge stacks should be vertical and terminate at a point
where height or air velocity limits re-entry into supply air
inlets or other plant openings (see Section 5.3).
5.2.6 Ancillary Equipment Design Considerations.
Provide duct supports of sufficient capacity to carry the weight
of the system plus the weight of the duct half filled with material and with no load placed on the connecting equipment at
the hood.(5.1,5.2) Where quick clamp systems are used, more
supports may be necessary.
Provide adequate clearance between ducts and ceilings,
dampers, explosion vents, etc., in accordance with the
National Fire Protection Association (NFPA) Codes and other
applicable standards and manufacturers’ instructions. Exhaust
fans handling explosive or flammable atmospheres require
special construction (see AMCA(5.3) for spark-resistant fan
construction guidelines). Consult NFPA and other sources for
correct specifications.
Minimize the use of blast gates or other dampers, if possible. However, if blast gates are used for system adjustment,
place each in a vertical section midway between the hood and
the next junction. To reduce tampering, provide a means of
locking dampers in place after the adjustments have been
made. Blast gates or orifice plates are mandatory if air balancing is required. Blast gates should be included in all ducts
where adjustment is required.
Allow for vibration and expansion. If no other considerations make it inadvisable, provide a flexible connection
between the duct and the fan. The fan housing and drive motor
should be mounted on a common base of sufficient weight to
dampen vibration, or on a properly designed vibration isolator.
Do not allow hoods and duct to be added to an existing
exhaust system unless specifically provided for in the original
design or unless the system design is modified. If changes are
made to the duct system, use methods shown in Chapter 8 of
the O&M Manual and Chapter 9 of this Manual. Locate fans
and filtration equipment to provide easy access for maintenance. Provide adequate lighting in penthouses and mechanical rooms.
Where federal, state, or local laws conflict with the preceding, the more stringent requirement should be followed.
Deviation from existing regulations may require approval by
local regulators.
5.3
DISCHARGE STACKS
The final component of the ventilation system is the exhaust
stack, an extension of the exhaust duct above the roof or grade.
Assuming all exhaust emission levels are met and maintained,
there are still two prime design considerations for the placement of an exhaust stack for a local exhaust ventilation system.
First, the air exhausted should escape the building envelope so
it does not return directly into building air intakes. Secondly,
once it has escaped the building envelope, the stack should
provide enough dispersion so that the plume does not cause an
unacceptable situation when it reaches the ground.
It is never recommended that the exhaust stack contain a
weather (or rain) cap. Such fittings do not successfully prevent
precipitation from entering the ventilation system and only add
to the energy requirement of the system. If precipitation and
ice are anticipated to present an issue in the exhaust stack or at
the fan exhaust/outlet, install appropriate drainage in the
exhaust stack and/or seek to utilize an offset stack or offset
elbows in the system design (see Figure 5-4). If the exhaust
stack design includes horizontal runs, the duct should be
slightly inclined toward a drain point.
Additionally, the fan should possess a drain port so that
moisture does not settle in its housing, potentially causing corrosion or drive system damage. Large, heavy, vertical exhaust
stacks should not be supported directly by the fan. Failing to
provide an adequate support structure for the discharge stack
may lead to early structural failure of the fan housing.
When placing an exhaust stack on the roof of a building, the
designer should consider several factors. The most important
is the pattern of the air as it flows over the building. Even in
the case of a simple building design with a perpendicular wind,
the airflow patterns over the building can be complex. Figure
5-1 shows the interaction between the building and the wind.
As air impacts the leading wall of a building, a downdraft is
created by deflected airflow. Additionally, a stagnation or
recirculation zone forms as air flows over the leading edge of
the building’s upwind wall. Vortices form by the wind action
as air rises and flows over the building’s leading edge, resulting in a roof recirculation region (see Z1 in Figure 5-1) along
the leading edge of the roof and/or roof obstructions. Further,
additional recirculation regions form as air flows over the
building’s roof and passes over the downwind edge of any roof
obstructions and the building’s edge.
The United States Environmental Protection Agency
(USEPA) uses computer modeling/simulations that employ
Gaussian distribution (such as PTMax) to predict resulting
ground level concentrations of pollutants emitted from stacks.
These predictive tools show 10-to-100 times the normal
ground level concentrations when building wake effects are
included; most frequently due to insufficient stack height.
More guidance in using these tools can be found at the
USEPA’s Support Center for Regulatory Atmospheric
Modeling (SCRAM).
The roof recirculation region forms in an area where a relatively fixed amount of air moves in a circular fashion with little air movement through the boundary. A stack discharging
into the recirculation zone can pollute the zone, resulting in
contaminated air entering the building through the HVAC system intakes located in the region. Consequently, all stacks
should penetrate the recirculation zone boundary.
The high turbulence region (see Z2 in Figure 5-1) is one
through which the air passes in a highly erratic manner with
significant downward flow. A stack that discharges into this
region will contaminate anything downwind of the stack.
Consequently, all discharge stacks should extend high enough
that the resulting plume does not intersect with the high turbu-
Duct System and Discharge Stack Design Principles
5-7
FIGURE 5-1. Effects of building on stack discharge
lence region, particularly upwind of a building air intake.
Because of the complex flow patterns around simple buildings, it is difficult to locate a stack that is not influenced by
vortices formed by the wind. Tall stacks are often used to
reduce the influence of turbulent flow, thereby releasing the
exhaust air above the influence of the building and preventing
contamination of the air intakes. Selection of the proper location is made more difficult when the facility has several supply
and exhaust systems, and when adjacent buildings or terrain
also cause turbulence around the facility. Additionally, prevailing winds should be considered in the design of any stack system, however, it should not be relied on to ensure efficacy of
the system. Wind rose plots can be used to aid in determining
the prevailing winds for various geographical locations.
The effect of wind on stack height varies with speed:
1) At low wind speeds, the exhaust jet from a vertical
stack will rise above the roof level resulting in significant dilution at the air intakes.
2) Increasing wind speed can decrease plume rise and
consequently decrease atmospheric dilution.
3) Increasing wind speed can increase turbulence and
consequently increase dilution, but this may trap diluted contaminants in the building’s wake zone.
Predicting the location and form of the recirculation cavity,
high turbulence region, and roof wake is difficult. However,
for wind perpendicular to a rectangular building, the height
(H) and the width (W) of the upwind building face determine
its airflow patterns. The critical dimensions are shown in
Figure 5-1. According to Wilson,(5.4) the critical dimensions
depend on a scaling coefficient (R) and are given by:
R = BS0.67 H BL0.33
[5.2]
where:
BS = the smaller of the upwind building face
dimensions H and W (ft) [m]
BL = the larger of the upwind building face
dimensions H and W (ft) [m]
When BL is larger than (8 × BS), use BL = 8(BS) to calculate
the scaling coefficient. For a building with a flat roof,
Wilson(5.4,5.5) estimated the maximum height (HC), center (XC),
and length (LC) of the recirculation region as follows:
HC = 0.22(R)
[5.3]
XC = 0.5(R)
[5.4]
LC = 0.9(R)
[5.5]
In addition, Wilson estimated the length of the building
wake recirculation region (LR) by:
LR = 1.0(R)
[5.6]
The exhaust air from a stack often not only possesses
upward momentum due to the exit velocity of the exhaust air,
5-8
Industrial Ventilation
but also buoyancy due to its density. The effective stack height
is used for the evaluation of the actual stack height (see Figure
5-2). The effective height is the sum of:
1) actual stack height (HS),
2) the rise due to the vertical momentum of the air, and
3) any wake downwash effect that may exist.
Note that these equations are valid only in IP units. Final
translation to metric units would be done after determining the
dimensions in IP units.
A wake downwash occurs when air passing a stack forms a
downwind vortex. This vortex draws the plume downward,
reducing the effective stack height (see Figure 5-3). The vortex
effect is eliminated when the exit velocity is greater than 1.5
times the wind velocity. If the exit velocity exceeds 3,000 fpm
[15.24 m/s], the momentum of the exhaust air reduces the
potential downwash effect.
The ideal discharge stack design extends high enough so
that the expanding plume does not meet the roof wake boundary (see Z3 in Figure 5-1). More realistically, the stack should
be extended so that the expanding plume does not intersect the
high turbulence region or any recirculation region. According
to Wilson,(5.5) the high turbulence region boundary (Z2) follows a 1:10 downward slope from the top of the recirculation
cavity (see Figure 5-1).
FIGURE 5-2. Effective stack height
To avoid entrainment of exhaust gas into the wake, a discharge stack must terminate above the recirculation cavity.(5.6)
The effective stack height to avoid excessive re-entry can be
calculated by assuming that the exhaust plume spreads from
the effective stack height with a slope of 1:5 (see Figure 5-1).
The first step is to raise the effective stack height until the
lower edge of the 1:5 sloping exhaust plume avoids contact
with all recirculation region boundaries. Note that recirculation regions may also be generated by roof top obstacles such
as air handling units, penthouses, or architectural screens. The
heights of the cavities are determined by Equations 5.2, 5.3,
and 5.4 using the scaling coefficient for the obstacle. Equation
5.5 can be used to determine the length of the roof recirculation region downwind of the obstacle.
If any air intakes, including windows and other openings,
are located on the downwind wall, the lower edge of the plume
with a downward slope of 1:5 should not intersect with the
building wake recirculation region downwind of the building.
The length of the building wake recirculation region (LR) is
given by Equation 5.6. If the air intakes are on the roof, the
downward plume should not intersect the high turbulence
region above the air intakes. When the intake is above the high
turbulence boundary, extend a line from the top of the intake
to the stack with a slope of 1:5. When the intake is below the
high turbulence region boundary, extend a vertical line to the
Duct System and Discharge Stack Design Principles
5-9
FIGURE 5-3. Wake downwash effects
boundary, then extend back to the stack with a slope of 1:5.
This allows the calculation of the minimum stack height; this
height can be determined for each air intake. The maximum of
these heights would be the minimum required stack height. In
addition, the heights may need to be increased to ensure that
plume does not intersect with the roof wake boundary, as discussed above.
In large buildings with many air intakes, the above procedure will result in the specification of tall stacks. An alternate
approach is to estimate the amount of dilution that is afforded
by stack height, distance between the stack and the air intake,
and internal dilution that occurs within the system itself. This
approach is presented in the “Airflow Around Buildings”
chapter in the Fundamentals volume of the ASHRAE
Handbook.(5.7)
In summary, the following should be considered for proper
stack design:
1) Discharge velocity and gas temperature influence the
effective stack height.
2) Wind can cause a downwash into the wake of the stack
reducing the effective stack height. Stack velocity
should be at least 1.5 times the wind velocity to prevent
downwash.
3) A preferable stack exit velocity is 3,000 fpm [15.24
m/s] because it prevents downwash for winds up to
2,000 fpm, or 22 mph [10.16 m/s]. Higher wind speeds
lead to increased dilution effects. Increased stack velocity also increases effective stack height and allows
selection of a smaller centrifugal exhaust fan. It can
also provide transport velocity if there is any particulate in the exhaust or if there is a failure of the air pollution control device.
4) High exit velocity is a poor substitute for stack height.
For example, a stack located at roof elevation requires
a velocity over 8,000 fpm [40.64 m/s] to penetrate the
roof recirculation region.
5) The terminal velocity of rain is about 2,000 fpm [10.16
m/s]. A stack velocity above 2,600 fpm [13.20 m/s]
should prevent rain from entering the stack when the
fan is operating.
6) Locate stacks on the highest roof of the building when
possible. If not possible, a higher stack is required to
extend beyond the wake of the high bay, penthouse, or
other obstacle.
7) The use of an architectural screen should be avoided.
The screen becomes an obstacle and the stack must be
raised to avoid the wake effect of the screen.
8) The best stack shape is a straight cylinder. If a drain is
required, a vertical stack head is preferred (see Figure
5-4). In addition, the fan should be provided with a
drain hole and the duct should be slightly sloped
toward the fan.
9) Weather (rain) caps should not be used (see Figure 55). Such caps direct the air toward the roof, increasing
5-10
Industrial Ventilation
the possibility of re-entry and potentially causing exposures to maintenance personnel on the roof.
5.2
Sheet Metal and Air Conditioning Contractors’
National Assoc., Inc.: Rectangular Industrial Duct
Construction Standards. Tysons Corner, Vienna, VA
(2011).
5.3
Air Movement and Control Association, Inc.: AMCA
Standard 210-74. Arlington Heights, IL (2005).
5.4
Wilson, D.J.: Flow Patterns Over Flat Roof Buildings
and Application to Exhaust Stack Design. ASHRAE
Transactions, 85:284–95 (1979).
5.5
Wilson, D.J.: Contamination of Air Intakes from Roof
Exhaust Vents. ASHRAE Transactions, 82:1024–38
(1976).
5.6
Clark, J.: The Design and Location of Building Inlets
and Outlets to Minimize Wind Effect and Building
Reentry. Journal of the American Industrial Hygiene
Society, 26:262 (1956).
5.7
American Society of Heating, Refrigerating and AirConditioning Engineers: 2001 Fundamentals Volume,
Section 16.1. ASHRAE, Atlanta, GA (2001).
5.8
American Welding Society: (AWS D1.1-72) Miami,
FL (2008).
10) Separating the exhaust points from the air intakes can
reduce the effect of re-entry by increasing dilution.
11) In some circumstances, several small exhaust systems
can be placed in a single manifold to provide internal
dilution thereby reducing re-entry.
12) A combined approach of vertical discharge, stack
height, remote air intakes, proper air pollution control
device, and internal dilution can be effective in reducing the consequences of re-entry.
A tall stack is not an adequate substitute for good emission
control. The reduction achieved by properly designed air pollution control devices can have a significant impact on the
potential for re-entry. (This may not apply to scrubber exhaust
because of moisture.)
REFERENCES
5.1
Sheet Metal and Air Conditioning Contractors’
National Assoc., Inc.: Round Industrial Duct
Construction Standards. Tysons Corner, Vienna, VA
(2017).
Duct System and Discharge Stack Design Principles
FIGURE 5-4. Stackhead design
FIGURE 5-5. Rain caps
5-11
5-12
Industrial Ventilation
Duct System and Discharge Stack Design Principles
5-13
5-14
Industrial Ventilation
Duct System and Discharge Stack Design Principles
5-15
5-16
Industrial Ventilation
Duct System and Discharge Stack Design Principles
5-17
5-18
Industrial Ventilation
Chapter 6
HOOD DESIGN
NOTE: Equations with notation followed by (IP) are designated for inch-pound system only; equations followed by (SI) are designated for
metric use only. If equation bears neither, then it applies to both systems.
6.1
6.2
6.3
6.4
6.5
6.6
INTRODUCTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-3
6.1.1 Local Exhaust Hoods as Compared
to Dilution . . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-3
6.1.2 Local Exhaust System Effectiveness . . . . . . . . .6-3
6.1.3 Design Goals . . . . . . . . . . . . . . . . . . . . . . . . . . .6-4
6.1.4 Wake Zones . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-4
6.1.5 Hood Types . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-4
ENCLOSING HOODS – INTRODUCTION . . . . . . . . .6-5
6.2.1 Airflow Requirements for Enclosing Hoods . . .6-5
6.2.2 Enclosing Hood Face Velocity . . . . . . . . . . . . . .6-5
6.2.3 Enclosing Hoods that Rely on Uniform
Flow To Protect Workers . . . . . . . . . . . . . . . . . .6-6
6.2.4 The Importance of Uniform Flow . . . . . . . . . . .6-6
6.2.5 Achieving Uniform Face Velocities in
Enclosing Hoods . . . . . . . . . . . . . . . . . . . . . . . .6-7
6.2.6 Effect of Supply Air on Uniformity of
Flows at the Hood Face . . . . . . . . . . . . . . . . . . .6-8
6.2.7 Large Spray Booth Airflow Patterns . . . . . . . . .6-8
6.2.8 Bench Top Enclosing Hood Airflow Patterns . .6-9
6.2.9 Steps for Designing a Uniform Flow
Enclosing Hood . . . . . . . . . . . . . . . . . . . . . . . . .6-9
TOTALLY ENCLOSING HOODS . . . . . . . . . . . . . . . .6-11
6.3.1 Issues in Common . . . . . . . . . . . . . . . . . . . . . .6-11
6.3.2 Extremely High Control (EHC) Total
Enclosures . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-12
6.3.3 Highly Effective Control (HEC) Total
Enclosures . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-12
6.3.4 High Control (HC) Total Enclosures . . . . . . . .6-12
6.3.5 Moderate Control (MC) Total Enclosures . . . .6-12
HOT PROCESSES IN ENCLOSING HOODS . . . . . .6-13
DOWNDRAFT OCCUPIED HOODS
(CLEAN ROOMS) . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-13
CAPTURING HOODS . . . . . . . . . . . . . . . . . . . . . . . . .6-13
6.6.1 Shapes of Capturing Hoods . . . . . . . . . . . . . . .6-14
6.6.2
6.6.3
6.6.4
6.6.5
Capture Velocity . . . . . . . . . . . . . . . . . . . . . . . .6-14
Effective Zone of Capturing Hoods . . . . . . . . .6-16
Capturing Hood Shape and Placement . . . . . .6-21
Use of Slots in Slot Plenum (Compound)
Hoods . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-21
6.6.6 Airflow Requirements for Compound
(Slotted) Hoods (Aspect Ratio < 0.2) . . . . . . .6-22
6.6.7 Airflow Requirements for Aspect Ratios
Greater Than 0.2 . . . . . . . . . . . . . . . . . . . . . . . .6-24
6.6.8 Critical Issues for Capturing Hood
Airflow Equations . . . . . . . . . . . . . . . . . . . . . .6-25
6.6.9 Push-Pull Hoods . . . . . . . . . . . . . . . . . . . . . . . .6-26
6.6.10 Compensating Air Hood . . . . . . . . . . . . . . . . .6-26
6.6.11 Downdraft Hoods . . . . . . . . . . . . . . . . . . . . . . .6-26
6.6.12 Receiving Hoods . . . . . . . . . . . . . . . . . . . . . . .6-26
6.6.13 Steps to Designing a Capture Hood . . . . . . . . .6-27
6.7 CHOOSING BETWEEN CAPTURE AND
ENCLOSING HOODS . . . . . . . . . . . . . . . . . . . . . . . . .6-28
6.8 ERGONOMIC CONSIDERATIONS FOR
DESIGN OF HOODS . . . . . . . . . . . . . . . . . . . . . . . . . .6-28
6.9 WORK PRACTICES . . . . . . . . . . . . . . . . . . . . . . . . . .6-29
6.10 MATERIAL HANDLING IN AND NEAR HOOD
WORKSTATIONS . . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-29
6.11 HOOD MAINTENANCE AND CLEANING . . . . . . .6-29
6.12 HOODS AND PERSONNEL FANS . . . . . . . . . . . . . . .6-29
6.13 VENTILATION OF RADIOACTIVE AND
HIGH TOXICITY PROCESSES . . . . . . . . . . . . . . . . .6-30
6.14 DETERMINING HOOD STATIC PRESSURE
LOSSES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-31
6.14.1 Static Pressure Losses for Simple Hoods. . . . .6-32
6.14.2 Pressure Loss in Compound Hoods. . . . . . . . .6-33
6.14.3 Coefficient of Entry and System Evaluation . .6-34
6.14.4 Determination of Ce . . . . . . . . . . . . . . . . . . . . .6-34
REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-35
6-2
Industrial Ventilation
Figure 6-1
Figure 6-2
Figure 6-3a
Figure 6-3b
Figure 6-4
Figure 6-5
Figure 6-6
Figure 6-7
Figure 6-8
Figure 6-8a
Figure 6-8b
Figure 6-9
Figure 6-10
Figure 6-11
Figure 6-12
Figure 6-13
Figure 6-14
Figure 6-15
Figure 6-16
Figure 6-17
Figure 6-18
Figure 6-19
Figure 6-20
Flow with no Crossdraft . . . . . . . . . . . . . . . . . . 6-4
Flow with Crossdraft . . . . . . . . . . . . . . . . . . . . .6-5
Flow into a Capturing Hood . . . . . . . . . . . . . . .6-5
Flow into an Enclosing Hood . . . . . . . . . . . . . .6-5
Parts of an Enclosing Hood . . . . . . . . . . . . . . . .6-7
Multiple Takeoffs for Wide Hoods . . . . . . . . . .6-8
Tapered Entry . . . . . . . . . . . . . . . . . . . . . . . . . . .6-8
Skewed Entry . . . . . . . . . . . . . . . . . . . . . . . . . . .6-8
Auxiliary Flow Hood . . . . . . . . . . . . . . . . . . . . .6-9
User-occupied Enclosing Hood
Recommendations . . . . . . . . . . . . . . . . . . . . . .6-10
Benchtop Enclosing Hood
Recommendations . . . . . . . . . . . . . . . . . . . . . .6-11
Ineffective Hot Process Hood . . . . . . . . . . . . .6-13
Enclosing Hood Designed for Hot Source . . .6-14
Downdraft Room . . . . . . . . . . . . . . . . . . . . . . .6-15
Plain Opening Hood . . . . . . . . . . . . . . . . . . . .6-15
Slot Hood . . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-15
Slot-Plenum Hood . . . . . . . . . . . . . . . . . . . . . .6-16
Flow/Capture Velocity – Suspended and
Canopy Hoods . . . . . . . . . . . . . . . . . . . . . . . . .6-17
Flow/Capture Velocity – Slotted Hoods . . . . .6-18
Flow/Capture Velocity – Downdraft and
Booth-Type Hoods . . . . . . . . . . . . . . . . . . . . . .6-19
Effective Capture Zone . . . . . . . . . . . . . . . . . .6-20
Flow Rate as Distance from Hood . . . . . . . . .6-20
Velocity Contours – Flanged Duct and
Plain Duct End . . . . . . . . . . . . . . . . . . . . . . . . .6-22
Figure 6-21
Figure 6-22
Figure 6-23
Figure 6-24
Multiple Slot Hood . . . . . . . . . . . . . . . . . . . . .6-22
Slot Hood with Baffles . . . . . . . . . . . . . . . . . .6-24
Buoyant Source and Horizontal Flow . . . . . . .6-24
Incline and Elevate Capturing Hoods for
Buoyant Sources . . . . . . . . . . . . . . . . . . . . . . .6-24
Figure 6-25 Plain Opening Acts as a Point Sink . . . . . . . . .6-25
Figure 6-26 Push-Pull Ventilation for Dip Tanks . . . . . . . .6-26
Figure 6-27 Compensating Air Hood . . . . . . . . . . . . . . . . .6-27
Figure 6-28 Downdraft Hood . . . . . . . . . . . . . . . . . . . . . . .6-27
Figure 6-29 Overhead Canopy Hoods . . . . . . . . . . . . . . . . .6-27
Figure 6-30 Small Enclosing Hood . . . . . . . . . . . . . . . . . . .6-29
Figure 6-31 Chain Slot . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-30
Figure 6-32 Roll Out Hood . . . . . . . . . . . . . . . . . . . . . . . . .6-30
Figure 6-33 Turntable . . . . . . . . . . . . . . . . . . . . . . . . . . . . .6-30
Figure 6-34 Dip Tank Hood That Drains Condensed
Fluid from Plenums . . . . . . . . . . . . . . . . . . . . .6-31
Figure 6-35 Hopper Bottom to Ease Removal of
Settled Materials . . . . . . . . . . . . . . . . . . . . . . .6-31
Figure 6-36 Separation of Flows at the Duct Inlet and
Hood Loss Factors; Values Shown
for Round Entries . . . . . . . . . . . . . . . . . . . . . . .6-31
Figure 6-37a Measurement Location for SPfilter in
Typical Enclosing Hood . . . . . . . . . . . . . . . . .6-32
Figure 6-37b Measurement Locations for SPfilter with
Filter at Entrance to Hood and at the
Plenum Face . . . . . . . . . . . . . . . . . . . . . . . . . .6-33
Figure 6-38 Compound Losses in Slot/Plenum Hood . . . .6-34
Figure 6-39 Hood Loss Factors . . . . . . . . . . . . . . . . . . . . . .6-36
____________________________________________________________
Table 6-1
Table 6-2
Abbreviations Used in Chapter . . . . . . . . . . . . . . .6-3
Recommended Capture Velocities . . . . . . . . . . . .6-21
Table 6-3
Summary of Hood Airflow Equations . . . . . . . . .6-23
Hood Design
6-3
TABLE 6-1. Abbreviations Used in Chapter
X
= greatest distance between contaminant and hood face
= slot area
Ce
= hood flow coefficient
= depth of hood or plenum
Fa
= acceleration (or Bernoulli) coefficient = 1
Hslot = height of hood, table, slot
Fh
= duct entry loss factor
L
= length of hood, table, slot
Fs
= slot loss factor
Q
= airflow requirement
SPh = hood static pressure
W
= width of hood, table
SPf
Vf
= hood face velocity
VPd = duct velocity pressure
Vx
= capture velocity necessary at distance X from
the hood face
VPs = slot or opening velocity pressure
Af
= face area of hood opening
As
D
6.1
INTRODUCTION
The starting point in the design of all industrial ventilation
systems is the interface between worker exposure to an air
contaminant and the Engineering Control (Industrial
Ventilation System) used to control that contaminant. It is the
point that determines whether successful contaminant control
can be assessed and therefore dictates the design of the rest of
the system. Chapter 1 speaks to the fact that all industrial
processes generate a dust or vapor cloud that becomes greater
in volume and less concentrated over time. Capture of that
generated contaminant cloud volume is dependent on several
complex variables that will be addressed in this chapter. The
human interface with an air contaminant may be the most
critical portion of industrial ventilation design. Not only is the
hood the most critical part of the design, but it is arguably the
most difficult.
The local exhaust hood is the point of entry into the exhaust
system and should include all suction openings regardless of
their physical configuration. It is the point that functions to
create an airflow field that will effectively capture the
contaminant-generated air and the air required to transport it
into the exhaust system. Design as a word describes applied
engineering skills, yet the interface with an occupational
airborne exposure deserves proper address by an Industrial
Hygiene practitioner. Hood design is where the industrial
hygienist and the engineer must overlap. Both disciplines must
be involved or the design will often be more or less than the
minimum amount of air required to protect the worker. This
Manual should be applied by persons that can incorporate both
disciplines.
If the hood design airflow is more than needed, it will put
the industry at an economic disadvantage as compared to the
same industrial process properly designed. Large systems
require larger contaminant control systems with larger motors
and larger amounts of heated and cooled replacement air (see
Chapter 11) for the life of a control system. That said, systems
designed with less air than that determined effective will cause
Wfl
= filter pressure loss
= flange width
worker exposure. A complete understanding of how the
employee interfaces with the air contaminant source must be
incorporated into all hood designs. All operator work practices
must be considered, as well.
This chapter will describe to the designer:
•
How to determine how much air volume will satisfy
the criteria,
•
The physics and aerodynamics of how to assess the
proper negative pressure requirements needed to capture the air contaminant and prevent exposure to a
worker,
•
The two main hood types along with hood nomenclature.
6.1.1 Local Exhaust Hoods as Compared to Dilution.
Ventilation hoods are the intake point of the local exhaust zone
that is created at the source of the contaminant before it has
time to be released to the ambient work air. Local exhaust
ventilation systems need only a fraction of the airflow required
for the best dilution ventilation system design (see Chapter
10). A concentrated source located near a worker is likely to
produce high exposures to that worker if only dilution
ventilation is employed.
6.1.2 Local Exhaust System Effectiveness. The ability of
a local exhaust ventilation system to reduce exposure to air
contaminants is determined primarily by three factors:
1) The hoods have sufficient exhaust airflow to contain
and capture contaminant,
2) The ability of the fan/duct system to deliver sufficient
airflow to each hood,
3) Work practices near the hood.
Chapter 13 contains recommendations for hoods for
specific processes and tasks based on the general principles in
this chapter. The designs can be adapted for different processes
and tasks, especially if the operating conditions and degree of
hazard are similar.
6-4
Industrial Ventilation
6.1.3 Design Goals. Exhaust airflow into a hood should
reduce the user’s air contaminant exposure to regulated levels
or lower. The levels of exposure are those required for
compliance with governmental regulations (e.g., OSHA, EPA)
or conformance with recommended practices (e.g., ACGIH®
TLVs®). Proper hood design should meet the following
requirements:
•
It should use the minimum airflow required to protect
workers
•
It should be designed with good ergonomic principles
in mind
•
It should be compatible with material flow through the
work area
•
It should allow for the inspection of internal parts of the
hood
•
It should be designed to minimize maintenance and
other activities that disrupt the process
6.1.4 Wake Zones. Understanding the wake zone is
important when designing or operating hoods. Air passing
around any blunt obstruction (including the human body)
creates a complex downstream counter-flow known as a wake
zone. This includes stable recirculating airflow patterns called
vortices as well as flow back towards the obstruction. Wake
zones must be considered when designing hoods.
If the contaminant is released within the wake zone
downstream of a human body it can circulate in that zone.
Gradually it will dissipate due to dilution and sudden
downstream movement of vortices. This is called shedding.
Meanwhile, the backflow can also carry contaminants released
several feet downstream of the obstruction back towards the
body and up to the breathing zone.
only gradually dissipate with time.
For enclosing hoods there are separation zones associated
with flows around the perimeter of the hood (Figures 6-1 and
6-2). Ordinarily, contaminant that reaches those zones
probably would not be a problem if the worker is centered on
the hood or if the contaminant never reaches the perimeter.
However, if a high-velocity crossdraft approaches the hood
from 90°, the size of the separation zone on that side may be
large enough to intersect the wake zone of the worker’s body
and transfer contaminants between the two zones.
6.1.5 Hood Types. Hoods may have a wide range of
physical configurations but can be grouped into two main
categories: enclosing and capturing (sometimes called
external). A capture hood handles contaminant in front of the
face of the opening (Figure 6-3a). If the contaminant is pushed
by moving air, thermal buoyancy, or the momentum from the
contaminant release towards the capturing hood, the capture
hood is called a receiving hood. Some capturing hoods protect
workers working very near them (e.g., welding hoods) and
others serve to reduce background concentrations (e.g., high
canopies over furnaces).
If the contaminant is released within the confines of a
ventilated structure it is called an enclosing hood (Figure 63b). With careful design, some enclosing hoods can be used
with workers inside.
Both types of hoods can be effective for cases where either
the contaminant generation rate and/or the amount of
dispersion are relatively low. If the contaminant generation
rate is very high and highly dispersed, then only enclosing
hoods are likely to be effective.
If the flow is from the side or front of the body, the wake
zone is on the opposite side or the back of the body. Since the
mouth and nose generally face towards the front, they are not
in the wake zone unless flow is from the back. Facing 90° to
the crossdraft may provide the lowest exposures, depending on
how the contaminant is released to the air. Facing upstream
often will produce lower exposures than facing downstream if
the contaminant cloud does not extend above waist height.
Even under those conditions, having the worker face upstream
is discouraged because seemingly minor changes in work
practices may cause at least some contaminant to be dispersed
well above waist height.
Although there may be larger blunt bodies with larger wake
zones than the human body, the human body is the most
important because the backflows in its wake zone can draw
contaminants to the user’s breathing zone. Separation of flows
from surfaces produces conditions with some similarities to
the wake zones downstream of blunt bodies. Anytime airflow
changes direction when flowing around a surface, its
momentum causes some degree of separation of the flow from
that surface. Contaminants released into a separation zone will
FIGURE 6-1. Flow with no cross draft
Hood Design
6-5
FIGURE 6-3b. Flow into an enclosing hood
FIGURE 6-2. Flow with cross draft
6.2
ENCLOSING HOODS – INTRODUCTION
Enclosing hoods are ventilated boxes completely or partially
enveloping contaminant generation points. Enclosing hoods
prevent the escape of contaminant by physically limiting the
openings through which contaminated air can escape.
In general, enclosing hoods are the most effective means of
contaminant control but can be limited by the necessary access
by workers (see Section 6.9 – Work Practices). Generally, the
smaller the total area of permanent openings, the less airflow
is required and the better the containment of the contaminants
inside the enclosure. However, the containment efficiency of
such hoods generally results in high concentrations inside the
hood, making them unsafe for worker entry.
If workers spend substantial durations reaching through
permanent openings to manipulate objects inside the
enclosure, then openings must be located efficiently and
conveniently.
Often such hoods are mounted on stands, cabinets, or tables
so that the opening extends from waist height to above the
head of the worker and are called bench top hoods. A
laboratory hood is a bench top workstation hood. Some
enclosing hoods are large enough for the worker to stand
inside. Protection of workers in such hoods depends on the
uniformity of flows down the length of the hood since lateral
and upstream flows will draw contaminants from downstream
sources to the workers’ positions. The most noteworthy
application of this design is for spray painting large objects, so
these hoods are sometimes called spray booths even when not
used for spray painting.
6.2.1 Airflow Requirements for Enclosing Hoods. The
system should deliver enough airflow to maintain the desired
target velocity at the face of the hood (Vf) over the area of the
hood face (Af). Thus, airflow rate (Q) is computed from:
Q = Vf Af
[6.1]
3
where: Q = airflow rate, acfm [am /s]
Vf = average velocity at the face, fpm [m/s]
Af = total open area at the hood face, ft2 [m2]
For example, if the open face is 10' × 15' [3.0 m × 4.6 m]
and the face velocity (Vf) is 100 fpm [0.51 m/s], then:
Q = (100 fpm)(10 ft)(15 ft) = 15,000 acfm
[Q = (0.51 m/s)(3.0 m)(4.6 m) = 7.04 am3/s]
To keep airflow rate (Q) to a minimum, the open area must
be kept to a minimum consistent with the requirements of the
process.
Note the airflow requirements for hoods are not affected by
density. It is the velocity into the hood that determines the
effectiveness of the hood, not the mass rate. For example, the
same hood used for the same purpose in New Orleans (Sea
Level) and Denver (5,000 ft above Sea Level) [1524 m] should
both have the same face velocity (e.g., 100 fpm [0.51 m/s]).
Note that airflow rate on all figures in Chapters 3, 9, and 13
are stated in acfm unless specified otherwise to make it clear
that one should not consider the altitude and temperature of the
air entering the hood when setting hood airflow requirements.
The airflow requirements must also maintain proper
conditions inside the hood. This includes LEL or other exposure
requirements, even if face velocity values are maintained.
FIGURE 6-3a. Flow into a capturing hood
6.2.2 Enclosing Hood Face Velocity. Air movement
prevents the escape of contaminated air through the open face
of the enclosure. Within limits, the higher the airflow (i.e.,
velocity) through the face, the less contaminant escapes. The
design face velocity (Vf) should be based on the effectiveness
(concentration outside/concentration inside) required to
6-6
Industrial Ventilation
protect workers. In general, the minimum acceptable face
velocity should be determined by the:
Toxicity of the contaminant – A higher hazard generally
requires a higher velocity, but no studies have established how
much more effectiveness is gained for additional increments of
velocity. It is likely but not clearly demonstrated that the gain
is very small when the velocity already exceeds 150 fpm [0.75
m/s]. This is probably also true for all of the following
determinants. USEPA Method 204 provides a requirement for
200 fpm [1.01 m/s] for containment of volatile organic
compounds (VOC) and may be required for certain other
applications (see Chapter 13, VS-75-40). This requirement has
also been applied by some regulatory agencies for control of
other materials.
Rate of generation of the contaminant – A higher generation
rate generally requires velocity closer to the top of the
recommended range.
velocity is already above 120 fpm [0.60 m/s], even substantial
increases in face velocity might reduce exposures only a small
amount. To obtain substantial improvements in effectiveness it
is likely to require changing the design of the hood, improving
the work practices of those using the hood, reducing
crossdrafts or reducing the rate of generation of airborne
contaminants within the hood. There may be extra regulatory
requirements for hood velocities, such as USEPA Method 204.
Consult those references for hood designs that are required to
meet those criteria. Some regulatory agencies may also set
requirements for the control of fugitive emissions from
sources. EPA Method 204 (Chapter 13, VS-75-40) describes
the requirements of a permanent or temporary total enclosure
from the EPA perspective. In general, a reduction in fugitive
emissions from a hood will be associated with reduced
ambient concentrations in the immediate area near the hood.
6.2.3 Enclosing Hoods that Rely on Uniform Flow to
Protect Workers. The two most successful hoods ever
Strength of competing air motions inside the enclosure (e.g.,
pneumatic spraying) and outside the hood (e.g., crossdrafts,
personnel cooling fans, passing vehicles) – Very strong
competing air motions may warrant face velocities above the
range typically recommended. It is also quite possible that for
very poor conditions, exposures simply cannot be controlled
sufficiently to protect a worker who is very close to the source.
Based on a study of laboratory hoods, one source(6.1)
recommends taking steps to reduce crossdrafts to no more than
half of the hood face velocity. However, it is likely that to
avoid having crossdrafts approach the hood from 90° is
equally important.
designed, paint-spray booths and laboratory hoods (see
Chapter 13) are both enclosing hoods. They are more effective
in protecting worker exposures of hazardous and toxic
materials. In many cases, work tasks require workers either to
stand or sit and reach into the enclosure or to work inside the
hood. The contaminant cloud inside the hood must be
prevented from reaching the breathing zone. This is best
accomplished by preventing the contaminant from mixing
with the wake zone of the worker as much as possible. In most
cases enclosing hoods make this exposure reduction possible
and prohibit the release of contaminants to the work
environment.
Degree of enclosure employed – If the source is poorly
enclosed, the face velocity must be higher to compensate for
the lack of shielding from competing air currents. It is likely
that increasing velocities cannot completely compensate for a
poor enclosure.
6.2.4 The Importance of Uniform Flow. In Chapter 3, this
Manual describes the three types of flow as found in the field
of fluid dynamics. It describes what happens to airflow as
laminar, transitional (semi-laminar) or turbulent. For the
purposes of this chapter, we find the term uniform flow is a
better term to use instead of transitional flow. As an example,
if a worker is at the face of the hood reaching into it to work,
then the inflowing air must push the contaminant towards the
back of the hood and the contaminant should not recirculate to
the hood face. The primary strategy is to provide relatively
uniform velocities at the face of the hood and well into the
enclosure. A flow that has a uniform velocity will show little
swirl (spiraling flow), no large-scale eddy currents (thus no
stagnation zones with rotating flow) and no flow back toward
the face, which can occur under turbulent flow conditions.
Obstructions and competing air movements tend to disrupt the
uniformity of the airflow and thus reduce the protection
provided by the hood.
Size of the hood used – A bench top hood is generally small
enough that the user blocks a substantial fraction of the
opening. It is likely that wake effects from air flowing over the
back are worsened by that blockage, though probably to a
lesser degree than for a person inside a booth.
For bench top hoods, hood effectiveness increases
significantly with face velocity within the range of 75 to 130
fpm [0.38 to 0.66 m/s] and sometimes higher. For cases where
crossdrafts or competing air motions near or at the source are
severe (e.g., pneumatic paint spraying), face velocities above
150 fpm [0.76 m/s] could be required unless the user stands
well away from the region of contaminant dispersion.
For occupied hoods with relatively undisturbed flow from
the face to the plenum (e.g., spray booths), a range of 100 to
150 fpm [0.51 to 0.76 m/s] is usually adequate for typical
applications and conditions. (See VS plates in Chapter 13 for
other values on special hoods.)
If a hood’s performance is not adequate and the face
The same issues apply to personnel who work inside a large
enclosing hood. It is important that the movement of air
separate their wake zones from the contaminant cloud. The
separation is best achieved by distance, uniformity of
velocities through the enclosure, and by keeping contaminant
clouds downstream of or to the side of workers. Obstructions
Hood Design
and competing air motions can disrupt the flow in ways that
move the cloud toward the workers inside the enclosure.
To better accomplish uniform airflow, hoods generally have
a completely open face that is the same cross-section as the
enclosure. For occupied hoods, the face can be a wall of filters
to remove room air dust, especially for paint-spray booths.
While not open, a cross-section of the filter wall equal to
enclosure size can provide relatively uniform flow.
If the hood face is partially blocked so that little or no flow
passes through substantial portions of the face, the result will
be large-scale eddy currents, stagnant zones and movement of
contaminated air. If workers are inside the hood such
blockages can increase their exposures. If the worker is at the
face of the hood face, blockages can draw contaminants
toward the face of the hood and also increase exposures. In the
case of a laboratory hood, the barrier is a sash that can be
raised and lowered or moved laterally. It is intended to keep
the user’s face out of the hood. A vertical sash also serves to
keep the worker’s face well above the bottom of the sash. Lab
hoods (see Chapter 13, Section 13.35) have non-uniform flow
inside the enclosure, but the sash protects users by keeping
their heads outside of the hood.
Every aspect of the design of such hoods is affected by the
need to maintain uniform airflow. If the contaminant is carried
by air currents back toward the user, the hood may provide
very poor protection.
6.2.5 Achieving Uniform Face Velocities in Enclosing
Hoods. The area of the face of most enclosing hoods designed
for frequent worker access is very large compared to the crosssectional area of the connecting duct. Air passing through a
FIGURE 6-4. Parts of an enclosing hood
6-7
hood face must converge to the much smaller area of the duct
while accelerating to the higher velocity in that duct. Even
without the effects of crossdrafts, the face velocity is not likely
to be uniform across the face and the velocity could be very
low at some points across the face. In those cases, the
contaminant might escape at weak points. To improve the
uniformity of the flow velocities at the face and inside the
hood:
1) Make the hood relatively deep by setting a minimum
enclosure depth (Dencl in Figure 6-4) of at least 0.75
times the face height (H) or face width (W) – whichever is greater. Even if the velocity at the back of the hood
(see the right side of Figure 6-4) is not uniform, an adequate depth will assist in providing a relatively uniform
velocity near the hood face.
2) Install a plenum. A plenum is a box that holds similar
pressure at all points within the box and therefore
allows equal airflow ingress at any point on the surface
of the plenum. It is typically the section at the back of
the hood formed by a wall of filters, baffles or slots
(Figure 6-4). Filters and baffles cause the air to spread
evenly at the back of the booth. If filters are used, the
static pressure drop across the filters when clean should
be > 0.10 "wg [25 Pa].
If baffles or slots are used instead of filters, the total
cross-sectional area of the baffles should be 90–95% of
the face area. If slots are used, there should be at least
three, and they should be spaced evenly over the
plenum face. Steel mesh, expanded metal, and perforated metal can be use instead of baffles, slots or filters.
6-8
Industrial Ventilation
Static pressures in the plenum typically range from -0.4
to -0.7 "wg when designed correctly. Slot velocities
must exceed twice that of any internal plenum velocities in order for even distribution to occur. Internal
plenum obstructions can only occur parallel to airflow,
never perpendicular to airflow, as this will cause nonuniform distribution of flow and pressure.
3) To guide the air converging from the plenum section,
design the transition to the duct or takeoff (Figure 6-4)
with an included angle of 90° (taper angle of 45°). If
vertical space is not sufficient for a 90° included angle
take-off, consider multiple take-offs across the width of
the plenum (Figure 6-5).
4) Install a rounded or tapered entry at the hood face with
a radius greater than 2" [50 mm] (see right side of
Figure 6-6) to reduce the separation zones that are
inside the hood at the perimeter of the face. If the hood
must be extremely effective and the contaminant may
be released near the sides or top, consider installing airfoils to the perimeter of the hood face (installing a sash
may be more effective).
5) The hood face should extend the full width and height
of the enclosure to reduce wake zones, where possible.
6.2.6 Effect of Supply Air on Uniformity of Flows at the
Hood Face. If the path of the supply air to the hood is at some
FIGURE 6-6. Tapered Entry
Even if its pathway is straight into the hood, supply air at
high velocity near the hood can be a problem. Excess airflow
can actually reverse course and exit through the face of the
hood, carrying contaminants with it. There is anecdotal
evidence that suggests that flow straight into a laboratory hood
could be more disruptive than air approaching at 90°.
The supply air should be delivered to the room through a
supply air duct system with its own fan and should be released
with a low initial momentum in the direction of the exhaust
hood but at a substantial distance from the hood (see Chapter
11, Section 11.3).
angle to the hood face, the airflow distribution at the face
might be skewed (Figure 6-7). The greater the velocity of the
crossdraft and the closer its angle is to 90°, the more disruptive
the supply air will be. In general, approach velocities should be
less than 30% of the hood face velocity.(6.2)
6.2.7 Large Spray Booth Hood Airflow Patterns. For
some operations (e.g., paint spraying) workers must occupy
the hood to do their work. To prevent transport of
contaminants towards the workers, hoods should be designed
to ensure that flow is relatively uniform and without backflow.
The flows should be aligned with the sidewalls all along the
length of the hood. Although this plug flow is effective in
carrying contaminated air away from the worker when the
source is downstream, they could produce more substantial
wake zones downstream of blunt bodies, including the
workers’ own bodies. These wake zones are likely to be much
more stagnant and larger than those seen in front of workers
FIGURE 6-5. Multiple takeoffs for wide hoods
FIGURE 6-7. Skewed entry
Hood Design
6-9
standing at the face of a bench top hood that has no sash. The
reason is that much of the air entering a bench top hood comes
from the perimeter and flows inward toward the center of the
hood, partially filling the wake zone in front of the user.
Air flowing through an occupied hood should be parallel to
the walls to avoid producing large eddy currents, especially if
both the worker and the source are within the same eddy or
wake. This is done by making the hood relatively deep and by
making the flow at the back of the hood as uniform as possible
by the use of panels, baffles or filters and by using 45° tapered
takeoffs. Large objects in the hood can also produce a stagnation zone upstream.
6.2.8 Bench Top Enclosing Hood Airflow Patterns. For
enclosing hoods small enough that the worker is stationed at
the face of the hood (i.e., bench top enclosing hoods, including
lab hoods), some of the air entering the hood must flow around
the user’s body to get into the face of the hood (Figure 6-1).
The air that flows around the operator’s body creates a wake
zone in front of the operator (see Section 6.1.4).
Performance is more vulnerable to conditions at the face of
the hood than to conditions at the back of the hood.(6.3) Moving
the source closer to the front of the hood will generally
increase contaminant concentrations at the face. Extending the
sides out past the operator’s position could also be detrimental
since the wake zone would be moved to the back of the user,
giving it a greater chance to interact with him.
Objects that serve to guide air smoothly into the hood will
reduce the size of the separation from the sides and top of a
plain enclosing hood. At high crossdraft velocities, the
addition of a flange will have marginal effect. The only
effective enhancement is a broad airfoil shape such as those
found on laboratory hoods (see Chapter 13, Section 13.35). A
deep airfoil or bevel at the bottom edge of the hood may
reduce the vena cava at the floor of the hood, but it also may
actually increase exposures if it pushes the worker away from
the hood face.
Eddy currents produced by the body of the operator
potentially can be reduced by directing 20–40% of the supply
air in front of his body (Figure 6-8 and Chapter 13, Section
13.35). This flow from the top would increase the flow
separation on the top of the inside of the hood. If the
contaminant is mostly near the floor of the hood, the net result
could be a reduction in exposure if this flow is not released in
excessive amounts or with excessive velocity. However, if the
contaminant is released with enough energy to reach the top of
the hood, this enlarged flow separation could pull the
contaminant to the user’s face, potentially increasing
exposures.
This auxiliary supply airflow is difficult to adjust properly,
and when it is adjusted poorly it might lead to higher
exposures. Because it may also add hot or cold air to the
workstation and require some filtration, it may not be practical
for most installations.
FIGURE 6-8. Auxiliary flow hood
Channeling or blowing air from the leading edge of the
floor of the hood is another approach. The upward flow might
increase dilution of the wake zone in front of the worker,
potentially reducing exposures. However, it is important once
again not to blow the air so high that it pushes the contaminant
up to the worker’s face and increases exposures.
6.2.9 Steps for Designing a Uniform Flow Enclosing
Hood. For hoods where workers must frequently reach into or
work inside (Figures 6-8a and 6-8b), the steps are:
1) Observe the operation through several cycles and question workers and maintenance personnel about access
needs, work practices, materials handling, emergency
conditions, and maintenance. Check the size of the
enclosure by watching the process and its operators.
2) On the side where operators must frequently reach into
the enclosure, install an opening (called a hood face) to
give operators the access they need. It is highly desirable to have only one side left open. Make sure the
open face gives the operators sufficient room to perform tasks.
3) For maintenance and operator tasks that are done no
more than a few times an hour, include additional openings that give access where needed, but cover with
doors or panels that are easy to open and close and cannot be removed during machine operation.
6-10
Industrial Ventilation
FIGURE 6-8a. User occupied enclosing hood recommendations
4) At points where it is necessary only to see inside the
enclosure, consider installing clear plastic or laminated
safety glass windows or doors.
ensure that the supply air enters the hood at low velocity. Choose a target face velocity for the hood (see
Section 6.2.2).
5) Provide light inside of the enclosure. Install the fixture
on the outside of the hood so that its light shines
through a plastic or laminated safety glass window and
does not produce glare. If a fixture must be inside the
enclosure, consider whether explosion proof fixtures
and wiring are required by code.
9) For extremely toxic materials, consider commercial
laboratory hoods or glove boxes and follow the manufacturer’s instructions.
6) Make the enclosure convenient. See Section 6.8 for
ergonomic considerations.
11) To avoid exhausting materials that might plug the duct
(rags, etc.), install perforated plate with small diameter
openings at the back of the hood instead of baffles.
Provide access for inspection and cleaning the screen.
For sections of duct that will very likely be coated by
sticky material or are otherwise likely to plug, consider
installing 5' [1.5 m] lengths of duct manufactured to be
easily removed and reinstalled (e.g., with built-in
clamp connections).
7) Force the air to flow evenly at the face of the hood so
that the face velocity is reasonably uniform. Likewise,
the plenum of the hood should be a wall of appropriate
filters, baffles, perforated panels, or other materials
with 5–10% openings. Consider multiple take-offs to
improve airflow patterns and better duct transition if
overhead space is insufficient for a 45° tapered takeoff
(Figure 6-5).
8) Particularly for laboratory hoods and large hoods,
10) To avoid product pickup, extend the width or height of
the hood so that the duct opening is a sufficient distance
away from the source.
12) After the hood is installed and periodically thereafter,
evaluate its performance both for ventilation effective-
Hood Design
6-11
FIGURE 6-8b. Benchtop enclosing hood recommendations
ness and worker acceptance. If either is unacceptable,
make revisions to meet all design and operational
goals. Install pressure gauges to monitor hood static
pressure and ΔP across filters.
6.3
TOTALLY ENCLOSING HOODS
Total enclosures have varying degrees of completeness in
physical or architectural enclosure and varying levels of
stringency in attempting to prevent contaminant escape
through whatever openings are in place.
6.3.1 Issues in Common. Most total enclosures will have
higher concentrations of contaminant inside the hood. The
concentrations are generally higher because their high
containment efficiency leads to use of relatively low airflows
compared to the generation rate of contaminants. Stagnation
zones due to eddies can produce large differences in
concentrations within the structure and hot spots of high
concentrations. Because of backflows and stagnation zones,
concentrations at openings can be relatively high, making
these hoods unsuitable for operations where workers
frequently reach into the hood openings.
In almost all total enclosures, at least some material
handling is typically done through openings that are kept
blocked by panels and doors. Total enclosures are often the
only hoods capable of adequately controlling sources that are
highly hazardous due to toxicity, rate of emissions, and
energetic dispersion. The amount of air continuously
exhausted from the hood must also exceed the amount of
contaminant produced by evaporation and other mechanisms,
including rapid displacement of the air in the hood due to rapid
inflow of materials and thermal expansion. The recommended
airflow could be specified (see Chapter 13) based on the
expected effluent level for specific applications but typically is
stated as a minimum velocity through ports and other openings
(see USEPA Method 204 as used for the determination of
hoods containing volatile compounds, VS-75-40). The inlet
velocity must also be high enough to overcome the momentum
of airflows that impact areas near the openings or baffles used
6-12
Industrial Ventilation
to block these flows. Intake velocities should be higher for hot
processes to overcome the buoyancy of air and gasses.
The inlet air ports and the exhaust port should be located to
ensure that stagnant regions do not develop, especially if
volatile materials are being contained. If the equipment inside
the hood allows, it is desirable to create uniform flow (see
Section 6.2.4). For cases where high velocity air movements
are created inside the enclosure by the process, it is important
to avoid placing ports where high velocity air can impact them.
Normally sources in the enclosure should be at least 4
equivalent diameters away from any opening (see Chapter 9,
Section 9.3.4 for definition of equivalent diameter). All
enclosures must be evacuated at a flow rate that will provide
operation at safe levels of LEL and other safety standards.
Assuming adequate levels of airflow and avoidance of
stagnant regions, the range of containment efficiency with
total enclosures of different types is strongly affected by the
care taken in minimizing opportunities for contaminants to
escape. For purposes of discussion, they are divided here into
the functional groups: Extremely High Control (EHC), Very
High Control (VHC), High Control (HC), and Moderately
High Control (MHC). The actual degree of control of each is
determined not only by initial design but also by installation
and operation. A hood designed for one contaminant and set of
conditions may fall short of requirements when another
material is to be contained or the generation rate or other
conditions are changed.
6.3.2 Extremely High Control (EHC) Total Enclosures.
Some processes are so hazardous that extreme care must be
taken in minimizing any escape from the hood. Examples are
the handling of radioactive dusts and gases, deadly bacteria
and viruses. To achieve EHC effectiveness, hoods must have a
high degree of enclosure and extreme care must be taken to
prevent escape through ports and openings.
The highest containment efficiencies within this group are
obtained by using ventilated boxes and by restricting access
when contaminants are inside the enclosure. Manual access
may be provided by manipulators inside the enclosure
controlled from outside the enclosure. The enclosure is opened
only after a substantial purge period and thorough internal
vacuuming of toxic dusts, viruses or bacteria. To ensure
constant dilution, the inlet ports should be numerous and small
with circuitous paths for exhaust flows. Special regulations
and standards should be consulted for these design
requirements.
doors in series. Even this arrangement will allow some transfer
of airborne contaminants to the room unless grilles are placed in
the airlock doors to provide continuous dilution of the chamber
between them. This glove port design could be implemented on
any enclosure design.
Proper air purge and glove seal design are important for
effective control. Manufacturer standards and regulatory
standards must be checked before usage and specification.
6.3.4 High Control (HC) Total Enclosures. Similar designs
to the glove box (HEC) can be used for other less toxic or
hazardous operations. For example, some sandblasting can be
done in small rooms with the operator standing outside the
room to manipulate the sand blast hose through gloved ports.
Because the seals in some cases may not be tight and because
the operation depends on effective work practices (i.e., waiting
for the enclosure to purge itself of dust before access),
exposure levels can sometimes be exceeded.
Ventilated storage cabinets are another form of HC
enclosure and can be designed with an exhaust port and
multiple grilles to allow entry of supply air into the cabinet.
The exhaust port and grilles should be positioned at each end
of the cabinet with the grilles placed to avoid stagnant zones
within the cabinet. A door to access the stored chemicals or gas
cylinders is a potential vulnerability for two reasons: 1) if it is
not shut, the control offered by the cabinet will be poor, and 2)
if a stored liquid spills or leaks from the storage vessel, the
fluid can seep underneath the door unless the vessel stands in
a bucket with sufficient volume to hold spilled liquids. If a gas
cylinder is stored in such a cabinet and developed a leak, the
resulting pressure could exceed the negative pressure in the
cabinet, allowing toxic gases to flow through the grille.
6.3.5 Moderate Control (MC) Total Enclosures. These
hoods are most common for non-toxic operations and general
ventilation of large sources. If the enclosure is relatively large
and the velocity through openings is relatively high (e.g., 150–
200 fpm) [approx. 0.75–1.00 m/s], it can provide reliable
control. They also generally provide the most reliable control
of very hot and large quantities of contaminated air. Examples
of these hoods are shown in Chapter 13, Section 13.27 (Hot
Processes). There are three critical points concerning these
hoods:
6.3.3 Highly Effective Control (HEC) Total Enclosures.
1) Even if large enough for operator entry, they are seldom designed or suitable for human occupancy. The
location of the air entry and exhaust points is important.
Generally, they should be designed for flow from one
end to another.
Somewhat lower protection is offered by glove box type
enclosures (see Chapter 13, VS-35-20), that utilize impregnable
gloves securely attached to internal ports. The operator inserts
his or her arms into the gloves and views the inside of the glove
box through a plastic glass or laminated safety glass window. In
most designs, adding or removing materials or equipment to or
from the glove box is done through an “airlock” of two small
2) Because they are often filled with large pieces of
process equipment (e.g., melting furnaces, etc.), it is
sometimes necessary to add additional inflow locations
to ensure that air flows through otherwise blocked
areas. In placing any opening, it is important that the
opening not be in line with a jet of contaminated air
issued within the enclosure.
Hood Design
6-13
3) To operate with very high effectiveness, all large openings must be opened only for short periods of time and
when the emissions are highly concentrated.
Note that the opening for supply air can be quite large yet
still be effective for large sources if uniform flow or near
uniform flow is established, and the inlet is far from the
workers’ breathing zones and is not used for worker access.
6.4
HOT PROCESSES IN ENCLOSING HOODS
Enclosures with small amounts of heat added (soldering and
welding) usually do not require special consideration for the
effects of buoyancy on calculations and design. However, if a
large area near the floor of the hood is heated to a high
temperature (e.g., > 300 F [approx. 150 C]), the inward
movement of air at an open vertical face of the hood may be
insufficient to move the heated air towards the back of the
hood. Instead, heated air may spill out of the opening near the
top of the face (Figure 6-9) since the upward velocity would
be at least as great as the inward velocity of the air flowing
through the hood face. The positive pressure exerted by the
buoyant force of the hot air can exceed the negative pressure
in the upper sections of the hood, allowing heated air and
effluents to leak from cracks and other openings near the top.
It is important that openings in the vertical faces be as close
to the bottom as possible and there be no permanent openings
near the top of the enclosure. Also locate the takeoff at or near
the top (Figure 6-11) so that the exhaust direction is aligned
with the buoyant air movement. Exhaust from front to rear
(Figure 6-10) is not recommended.
In addition, considerations must be made for the creation of
hot gasses by the process inside the enclosure and the decrease
in density (and therefore the increase of volume) as air is
heated inside the enclosure. See Chapter 13, Section 13.27 for
a comprehensive discussion of controls for heated processes.
6.5
DOWNDRAFT OCCUPIED HOODS
(CLEANROOMS)
Downdraft occupied hoods are enclosures designed to have
a uniform flow that is vertical instead of horizontal. Downdraft
hoods that rely on uniform flow (see Section 6.2.4) to protect
the worker or to minimize unwanted dispersion of the
contaminants generally should be designed to deliver airflow
uniformly through the ceiling face and removed uniformly
from the floor (Figure 6-11). Downdraft designs have an
advantage over horizontal flow in that wake zones from the
worker and objects on the floor are mostly under the floor. The
direction of flow is almost always downward.
Non-uniform release of the supply air generally will not
produce uniform flow. Zero and low velocity regions will be
marked by very large eddies in stagnation zones. High velocity
releases of supply air can produce flows with sufficient
momentum to rebound from large items on the floor to the
FIGURE 6-9. Ineffective hot process hood
position of the worker. Large scale eddies also will transfer
contaminants laterally, making it difficult to separate the
worker from contaminant clouds. For that reason, airflow
should be released as uniformly as possible from the entire
area of the ceiling (see Chapter 13, Section 13.10).
It is still important to exhaust air from the room uniformly.
Exhausting from a limited region in the floor will also produce
stagnant zones in the non-exhausted area. Any contaminant
reaching those zones will only be diluted slowly, potentially
producing high exposures to workers standing in them.
6.6
CAPTURING HOODS
Capturing hoods do not enclose the source, but instead rely
on a flow of air into the hood opening to capture the
contaminated air. Note that the air converging on an exhaust
point accelerates more rapidly as it approaches the hood face.
As a result, hood effectiveness in capturing the contaminated
air improves rapidly with decreasing distances from the hood
opening. The effectiveness falls off sharply at distances far
enough from the hood face that the inward velocity is not
significantly greater than the competing velocities induced by
6-14
Industrial Ventilation
FIGURE 6-10. Enclosing hood designed for hot source
traffic, man-cooling fans, process machinery or other
influences. The higher the velocity and the less the
competition from outside air currents, the more contaminant
will be collected and the more effective the hood.
6.6.1 Shapes of Capturing Hoods. Capturing hoods can be
shaped many different ways to fit specific geometric
constraints and needs, but the main types are:
1) Plain opening (Figure 6-12): Hood with a round opening or a rectangular opening with H/W (height/width)
> 0.2. The open face can remain a fixed cross-sectional
area for some distance or immediately converge to fit
the duct.
2) Slot hood (Figure 6-13): Hood with a relatively narrow
slot height (H) compared to its length (Lslot) followed
by a straight or converging transition to the duct. An
opening with H/Lslot < 0.2 is classified as a slot.
However, it should be understood that the airflow
behavior actually changes gradually with changes in
the aspect ratio.
3) Slot hood with plenum (Figure 6-14): Hood with one or
more relatively narrow openings followed by a sudden
expansion into a plenum. Airflow characteristics in
front of the hood are similar to a flanged slot with no
plenum. Note that each of these hoods has a tapered
transition from the hood face down to the duct size. The
tapering has no effect on airflow but does affect static
pressure requirements.
6.6.2 Capture Velocity. The minimum hood induced air
velocity necessary to capture and convey contaminant into the
hood is referred to as capture velocity (Figures 6-15, 6-16 and
6-17). In general, the effectiveness of capturing hoods
increases with increasing airflow levels and therefore with
increasing capture velocities (Vx). It is probable that an
increase in capture velocity can also offset the effects of
competing air currents, buoyancy, and contaminant
momentum. Therefore, higher capture velocities should be
used for higher crossdraft velocities. For buoyant plumes, it
may be more effective to place the hood above the level of the
Hood Design
6-15
FIGURE 6-11. Downdraft room
source. However, capture velocities are conceptually-based
and are not truly field verifiable because velocities degrade
evenly in all directions (see Figures 6-18 and 6-19) making
velocity measurements inconsistent.
Table 6-2(6.4,6.5,6.6) provides ranges of recommended capture
velocities for each of several examples with increasing
energies that serve to disperse the contaminated air. The ranges
are quite broad for each described dispersion condition. The
higher end of the ranges should be used for unfavorable
conditions, such as:
•
High crossdraft velocities,
FIGURE 6-12. Plain opening hood
•
Strong competing air motions due to traffic, mechanical motions, etc.,
•
Toxicity of contaminant, its generation rate, and the
duration of potential exposures.
The capture velocity should be at least 75 fpm [0.38 m/s]
except under ideal conditions. A velocity of 100 fpm [0.51
m/s] may be a more realistic minimum for typical conditions
(moderate toxicity, crossdrafts, etc.). It should be noted that a
capture velocity can also be excessive for some conditions. In
particular, very high capture velocities near dusty materials
can cause “product pickup.” The problem is most likely to
FIGURE 6-13. Slot hood
6-16
Industrial Ventilation
6.6.3 Effective Zone of Capturing Hoods. The effective
zone of a capturing hood is the region in front of the hood that
is adequately controlled by the flow of air into the hood
(Figure 6-18). The boundary of the effective zone can coincide
with the boundary where the induced velocity into the hood
equals the recommended capture velocity (Vx). However, the
two boundaries may be distinctly different if the contaminant
is highly buoyant, has its own momentum, or if there are
disturbing airflows due to crossdrafts, mechanical movement,
traffic, etc.
It is difficult to see air cross currents and to ascertain
whether contaminants are within the effective zone. The shape
and extent of the effective zone is affected by the exhaust flow
rate, the shape of the hood, nearby surfaces, crossdrafts, and
potential convection from hot sources. If the contaminant is
toxic or its generation rate is high, the hood efficiency must be
increased. The efficiency of capturing hoods is affected by the
following factors:
1) Distance from the source – inflow velocity decreases
dramatically with increasing distance from the hood
face (Figures 6-18, 6-19 and 6-20).
FIGURE 6-14. Slot-plenum hood
occur when the airflow through the hood is relatively low and
the hood must be kept very close to the source for the capture
velocity to be high enough to be effective. Using Figure 6-14,
1 ft2 hood [0.03 m2] with a capture velocity of 100 fpm [0.51
m/s] at 12" distance [0.3 m], the velocity at 6" [150 mm]
would be roughly 310 fpm [1.55 m/s]. That velocity could
cause entrainment of powdery products such as flour or talc. A
better solution is to enclose the source and make the hood
height sufficient enough that the region of high velocities near
the duct entry is far from the product (see Chapter 13, VS-5021 and VS-50-22). Not only would product pickup be
eliminated but the required airflow for acceptable performance
generally would be substantially lower than would have been
required for the capturing hood.
EXAMPLE PROBLEM 6-1 (Capture Velocity)
Determine capture velocity welding on mild steel, moderate
production, good conditions. The work table is 3' H 3' [.91 m H
.91 m].
Solution: (from Table 6-2) (average motion) Vx = 100–200
fpm [0.51–1.02 m/s]. Based on the stated conditions, the low
end of the range should be adequate, Vx = 100 fpm [0.51 m/s].
2) Location of the source – the source should be centered
immediately in front of the hood.
3) Shape of the hood – for some distance in front of the
hood, the velocity profile will differ depending on
whether the hood face is a slot or a plain opening. Long
slots produce velocity profiles that extend straight out
farther from the opening than do plain openings. The
effective zone of a plain hood will tend to be greater in
the vertical plane. (Note that the primary purpose of the
slot is to provide distribution over the length of the
slot.) As the distance from the hood face becomes
greater, all hoods begin to exhibit the performance profile of a plain hood. At near distances, a narrow slot
produces a cylindrical velocity contour for some distance in front of the slot. The slot opening is more
effective at distances on the same level as the slot, but
less effective for vertical distance above and below the
level of the slot. Using two or more parallel horizontal
slots, as in Figure 6-21, increases the efficiency of the
hood in the vertical plane. However, unless the slots are
relatively far apart (e.g., more than the desired effective
zone in the horizontal direction), the two slots will
behave more like a single large rectangular opening
than two single slot openings.
The relationship between capture velocity and airflow
(Q) for several hood shapes is shown in Figures 6-15,
6-16 and 6-17 and Table 6-3. Note that the source is
assumed to be directly in front of the hood opening.
4) Presence of surfaces near the hood that do not block the
flow – depending on their placement, such surfaces
may channel more of the airflow over the source, reducing the required airflow. For example, a flange partially
blocks the flow from behind the opening, increasing the
Hood Design
6-17
6-18
Industrial Ventilation
Hood Design
6-19
6-20
Industrial Ventilation
FIGURE 6-18. Effective capture zone
velocities in front of the hood. Likewise, resting the
hood on a tabletop can reduce the exhaust airflow
requirement because the airflow is channeled into the
hood, and side baffles also can channel airflow to the
hood face, reducing the exhaust airflow requirement.
Baffles perpendicular to the hood opening are sometimes used to block crossdrafts (Figure 6-22). They can
channel air over the source and into the hood opening
if crossdraft velocities are low. However, it is possible
that if crossdraft velocities are high, the upstream baffle
will create a strong wake zone that could reduce the
effectiveness of the hood rather than enhance it.
5) Objects and surfaces that impede flow across the
source and into the hood face – an object placed
between the source and the hood can channel the airflow so that it misses the contaminant.
6) Competing air currents – a high velocity crossdraft
(e.g., greater than 25% of the capture velocity) may
substantially distort the effective zone unless it is
blocked by other surfaces or objects. Likewise, competing air currents near the hood due to open windows,
personnel fans, mechanical or operator movements,
etc. can also distort and shrink the effective zone.
7) Motion of the contaminant – if the contaminant is released
at high velocity, it may fly away from the hood despite the
flow of air into the hood. In addition to the particle velocity, a competing air current has been created.
FIGURE 6-19. Flow rate as distance from hood
8) Buoyancy of the contaminated air – if the contaminated
air is rising rapidly because it is much warmer than room
air, its path becomes a complex function of the velocity
Hood Design
6-21
TABLE 6-2. Recommended Capture Velocities*
components in each direction induced by the air drawn
into the hood face and the upward velocity of the buoyant air. If the hood is drawing air solely in the horizontal
plane, the buoyant air may escape capture (Figure 6-23).
In those cases, the hood should be placed above the
source with its face angled approximately 45 degrees
with the vertical plane as is shown in Figure 6-24.
6.6.4 Capturing Hood Shape and Placement. The effects
of crossdrafts and other disturbances on the effective zone
should also be considered.
In general, a capturing hood should be at least 50% wider
than the anticipated diameter of the contaminant cloud. It also
should be at least as wide as the distance “X” (indicating the
greatest distance of contaminant from the hood face) (Figure
6-21). If the source can be placed anywhere on a work bench,
the width of the hood should be equal to the bench width if
possible. For example, if the source is constrained to be within
a 2 ft [0.6 m] width on the work bench, and the cloud of
released contaminants is less than 2 ft wide [0.6 m], and the
value of “X” is less than 2 ft [0.6 m], then the width of the
hood face should be 3 ft [0.9 m] (i.e., 50% wider).
The height needed for a capturing hood depends on the type
of hood, vertical height of the bulk of the emissions, and
buoyancy or upward momentum of the contaminated air. For
the same exhaust airflow, the height of the effective zone for a
hood with horizontal slots will be smaller than for a plain
hood. If the source is dispersed or rising over a significant
vertical distance, more than one slot may be required. If a slot
plenum hood is used, the plenum must extend high enough to
accommodate both openings. Note that if slots are located
close together (e.g., distance between midlines of slots is less
than the distance X), their effective zones will merge and it
will act like a plain hood.
For a plain hood, the effective zone vertically will be
roughly proportional to the vertical size of the opening for a
given exhaust volume. Greater exhaust volumes proportionally increase the effective size vertically and horizontally if the
source is relatively close to the hood.
The hood should be centered on the contaminant cloud if
the contaminated air is at room temperature and has no
significant momentum. If the source rests on a table top or
other work surface, the hood can be placed somewhat above
the emissions cloud, especially if the flange touches the table.
If the contaminated air is buoyant or has upward momentum,
the hood should be placed above and as close to it as possible
without interfering with the work (Figure 6-24).
6.6.5 Use of Slots in Slot Plenum (Compound) Hoods.
The primary reason to employ slots in a hood face is to force
uniformity of flow along the length of the slot. The length of
the slot should be greater than the width of the source in front
of it and its length also should increase with increasing
distance of the source from the hood face.
Some compound hoods have more than one slot, each
parallel to the long side of the hood. Since the effective zone
of a slot hood is limited above and below the plane of the slot,
then to ventilate sources at two heights, a slot should be placed
at each of the two heights (Figure 6-21).
For a given plenum, the higher the velocity through the slot
(Vs), the more uniform the velocities down the length of the
slot and the more uniform the flow in front of the hood.
However, slot velocities that are too high can cause hot spots
and uneven distribution. Since the airflow requirement is
determined based on other factors and the length is determined
by geometry, the velocity through the slot(s) can be influenced
only by setting the slot height (i.e., the smaller dimension
Hslot).
6-22
Industrial Ventilation
FIGURE 6-20. Velocity contours – flanged duct and plain duct end
The relatively low value of Vs = 1,000 fpm [approx. 5 m/s]
can produce adequate uniformity if:
1) The plenum has a velocity less than half the slot velocity,
2) The takeoff to the duct is centered on the slots and perpendicular to the slots (i.e., air makes a 90° turn after
entering through the slots) or the plenum is very deep
(e.g., depth = slot length),
3) The transition has a 45° or less taper angle, and
4) The closest slot is at least 1/2 of the slot length distance
from the taper.
If these conditions are not met, the slot velocity should be
higher. Velocities above 2,000 fpm [approx. 10 m/s] should
not be used as they are only marginally more effective.
If it is necessary to use an undersized plenum or if the takeoff will be at one end of the slots rather than centered on the
slots, it is likely that velocities down the length of the slot will
be progressively higher as the takeoff is approached even if Vs
= 2,000 fpm [10 m/s].
6.6.6 Airflow Requirements for Compound (Slotted)
Hoods (Aspect Ratio < 0.2). For hoods having an aspect ratio
FIGURE 6-21. Multiple slot hood
(height divided by length (H/Lslot) of 0.2 or less), only a small
fraction of the air flows from the ends (Figure 6-21) into the
Hood Design
TABLE 6-3. Summary of Hood Airflow Equations
6-23
6-24
Industrial Ventilation
FIGURE 6-24. Incline and elevate capturing hoods for
buoyant sources
FIGURE 6-22. Slot hood with baffles
face, so the air behaves to a large degree as if it were flowing
into a line sink. Therefore, at a distance X for a slot of length
L the control volume would be a cylindrical shape with a
surface area of A = 2πr  L. Since Q = V A, the airflow (Q)
required at a given distance would be Q = 2πr  L V and would
fall linearly with distance from the hood (Figure 6-16). The
approximate airflow (Q) required to achieve a specific velocity
(Vx) at a distance X for a slot of length L (upstream of the
midpoint of a freely suspended slot with no flange and with no
nearby obstructions) (Figure 6-16) is:
Q = 3.7 Vx L X
[6.2]
For compound hoods, if the slot is in the center of a large
flange, air is prevented from flowing from behind the hood,
thus improving its effectiveness in front. This can reduce
airflow requirements by as much as 20% for slots with aspect
ratios (H/L) equal to 0.25 and 35% for slots with aspect ratios
equal to 16.(6.7) For a flange width (Wfl ) greater than the square
root of the hood face area (i.e., Wfl /Af), a reasonable approximation is a reduction of 25% from Equation 6.2:
Q = 2.6 V × L X
[6.3]
If the slot is in a large wall (e.g., is cut into the plenum of a
compound hood), the airflow requirement should be even
lower than predicted by Equation 6.3. The maximum possible
reduction is 50% of the levels predicted in Equation 6.2. Other
surfaces near the hood can also reduce the airflow requirement
by channeling the air through the source and to the hood
opening. The most important example is the surface of a table
when a hood is on its surface.
Increasing slot velocity (by reducing slot height) while
holding Q constant will not extend the effective zone of the
compound hood. The total airflow requirement for a
compound hood with multiple slots is the sum of the
requirements for each of the slots. If the slots are the same size
and the plenum is of adequate size, the airflow through each
slot should be the same. When the slots are less than 0.5 H
apart they will act as a plain opening. Note also that tapering
from the hood face down to the duct has little or no effect on
airflow requirements.
6.6.7 Airflow Requirements for Aspect Ratios Greater
Than 0.2. The simplest possible hood would be a free standing
FIGURE 6-23. Buoyant source and horizontal flow
exhaust point. If we ignore the duct, the hood acts as a point
Hood Design
sink (Figure 6-25). In the absence of disturbing air currents,
the airflow would move toward the point sink uniformly from
all directions. At any distance X from the exhaust point, the
control volume would be a sphere with radius X and a surface
area of (4π)(X2). The mean velocity through the surface of the
imaginary sphere would be Q/Asphere. Thus, to establish any
given velocity Vx at a distance of X the required airflow would
be Q = Vx(4π)(X2). It can be shown that if the hood has an
aspect ratio greater than 0.2 or is round, then a hood in free
space with no nearby obstructions requires the airflow rate to
be (with Vx at distance X) estimated by:(6.8)
Q = Vx(10X2 + Af)
[6.4]
where: Af = area of face opening
However, capturing hoods often have relatively large
flanges that serve to block flow from the back of the hood,
increasing the flow from the front (Figure 6-20). For a flanged
hood in unobstructed space with Wfl /Af, the required airflow
is reduced:
Q = 0.75Vx(10X2 + Af)
[6.5]
Capturing hoods are often resting on a surface such as a
table top. If the hood rests on the table, the airflow requirement
reduces to:
Q = Vx(5X2 + Af)
[6.6]
And if the hood rests on the table and is flanged, the airflow
requirement reduces further to:
Q = 0.75Vx(5X2 + Af)
[6.7]
Since the contaminant from even very small sources may be
dispersed over a vertical height of several inches, it is usually
not advisable to place the hood directly on the work surface
6-25
unless it has a large flange resting on that surface. If the
contaminant is buoyant, the hood should be elevated above the
work surface (e.g., 12 to 24" [0.3 to 0.6 m]), with the height
increasing to a point with increasing thermal rise velocity. If
that is done, the airflow requirement should be somewhere
between Equations 6.5 and 6.6 since the work surface still
channels airflow to some degree but not as much as when
closer to the surface.
For both slot/plenum (compound) hoods and round or
rectangular shaped hoods, distance (X) is crucial. For
example, a 4" × 9" [101 mm × 229 mm] flanged hood that
draws 206 acfm [0.097 am3/s] will induce a velocity of 100
fpm [0.508 m/s] at a distance of 6 inches [0.152 mm], but only
27 fpm [0.137 m/s] at a distance of 12 inches [305 mm]. Any
measure that reduces the distance between hood face and the
source is likely to improve the performance of the hood.
The hood airflow equations are summarized in Table 6-3.
6.6.8 Critical Issues for Capturing Hood Airflow Equations. Equations 6.2 through 6.6 are based on the velocity
perpendicular to the hood face at the midpoint of the face for
ideal conditions. There is little research at this writing that can
be used to determine if the current recommendations are
optimal. It also should be noted that:
1) The equations model the velocity along the centerline
of the hood face, not at other points in the expected
control region. Real sources release contaminants that
may be spread over a substantial lateral range. Capture
velocity at the same distance from the hood face but not
at the centerline will be increasingly lower than the
midline velocity, especially for square and round hood
faces.
2) The equations do not consider the effects of crossdraft
velocities. It is reasonable to assume that the value of
capture velocity required to obtain the same effectiveness would increase substantially with higher crossdraft velocities.
3) The equations do not consider the effects of the worker’s body or the effects of work items placed between
the source and the hood.
4) The equations do not consider the effects of convection
air currents due to hot surfaces or effluents (e.g., welding plume) nor the effects of competing air currents due
to mechanical motions (e.g., spinning grinding wheel).
5) The equations for low aspect ratio hood openings probably apply much better to slot/plenum openings than to
slots that are not the open face of a plenum.
6) Two slots that are relatively close together (e.g., distance between them less than 0.5 H) will behave more
like a plain opening than a slot opening.
FIGURE 6-25. Plain opening acts as a point sink
7) The equations could overestimate airflow requirements
when the distance from the hood face exceeds 1.5 times
6-26
Industrial Ventilation
FIGURE 6-26. Push-pull ventilation for dip tanks
the hydraulic diameter (i.e., 4 times the area of the hood
face divided by its perimeter) of the hood face.
6.6.9 Push-Pull Hoods. Air emerging at high velocity from
a duct or nozzle can travel 30 diameters before turbulence and
expansion reduce its velocity to less than 10% of its initial
value. Air drawn into the face of a hood will have a velocity of
less than 10% of the face velocity at a distance of as little as
one duct diameter. Push-Pull systems (Figure 6-26) take
advantage of this by containing and pushing contaminated air
towards the capturing hood. Airflow reductions are possible
with short push distances but can be quite substantial for large
distances. See Chapter 13, Section 13.72 for detailed
descriptions and formulae for Push-Pull hood systems. A large
obstruction can reflect the push air away from the capturing
hood, especially if it is very close to the push jets (Figure 626). Air generally will flow around a moderate sized object,
especially if relatively far from the jets (e.g., more than five
times the smaller cross-sectional dimension of the
obstruction).
6.6.10 Compensating Air Hood. Another type of hood
blows clean air at low velocities at or near a capturing hood to
improve its effectiveness (Figure 6-27). This strategy can be
more effective than the capturing hood alone if applied
correctly. This design is effective with low velocity crossdrafts
(e.g., < 35 fpm [0.17 m/s]) and the supply airflow rate should
be adjusted to avoid blowing air past the hood. The exhaust
airflow rate should be at least 30% larger than the supply
airflow rate, and the release velocity of the supply air should
be less than 50 fpm [0.25 m/s]. These types of hoods have been
used successfully in foundries on shakeout and pouring sidedraft designs as well as high canopies. This is a new and
burgeoning field of design and effective distances of over 100
feet in length have been demonstrated in the field.
6.6.11 Downdraft Hoods. A downdraft hood is a type of
capturing hood with the air flowing downwards through a
horizontal face into the hood body (Figure 6-28). The
advantage of a downdraft hood is that large particles will fall
down through the grille covering the face to be collected in
cleanout drawers. Required airflow can be lower since the
distance to the source is reduced. As a capturing hood, the
necessary airflow can be computed using Equation 6.6 where
X is the maximum distance above the hood face where the
contaminant will be released.
It is important to note that X can be much higher than the
maximum height of the source if the work disturbs the air. For
example, a hand-held grinding wheel agitates the air by the
high rotation rate grinding surface.
It is important to recognize that operators can lay materials
or tools over the grille, blocking the airflow where it is needed
most, possibly rendering the hood useless.
6.6.12 Receiving Hoods. Receiving hoods are capture
hoods positioned so that: 1) particulate contaminants are
thrown into the hood opening from a distance, or 2) gas or
vapor contaminants are lifted by convection (buoyancy)
towards the hood opening. Overhead canopy hoods (Figures
6-15 and 6-29) are typically used to receive contaminants
mixed with heated air. Use of canopy hoods for very hot
processes (e.g., as found in work with molten metal) is
discussed in Chapter 13, Section 13.27.
Overhead canopy hoods are less effective for both warm
and ambient temperature air because:
1) Distribution of airflow can be poor if there are no positive measures taken to distribute air evenly at the hood
face. Air will flow preferentially near the top of the
face, not near the source where it may be needed most.
2) The open faces of this hood are the planes formed by
the perimeter of the source and the perimeter of the
Hood Design
6-27
FIGURE 6-28. Downdraft hood
atively low. The goal is to direct airflow over the contaminated source and into the hood.
3) Design hood face large enough to minimize frequent
movement of the hood.
FIGURE 6-27. Compensating air hood
canopy. Airflow enters from all four sides, so the air
volume requirements are correspondingly very large:
Q = (1.4)(tank perimeter)(X – height above
tank)(Vx)
[6.8]
3) If workers are near the source, contaminated air may be
directed into their breathing zone.
4) Fix the hood and the source in place if possible so that
the distance from it to the farthest point of contaminant
generation is always within the hood’s effective range.
5) To determine airflow (Q) requirements, first determine
the capture velocity needed considering the crossdrafts,
the toxicity of the contaminant, and the amount of the
contaminant. Recommended capture velocities are
shown in Table 6-2.
6) When necessary, consider covering the face of the hood
with a perforated plate to avoid picking up papers,
caps, rags, etc.
4) Since all sides are open, the hood is vulnerable to crossdrafts from all four sides.
The canopy hood can be improved by adding one to three
sides to it, but the distribution of velocities at the remaining
face will not be good (see distribution for enclosures, Section
6.2.5) unless baffling is added to the canopy face.
6.6.13 Steps to Designing a Capture Hood. When
designing a capture hood and selecting the airflow, consider
that crucial to its effectiveness is making sure the distance (X)
between the open face of the hood and the greatest distance to
a point of contaminant generation be kept as low as possible.
The steps to follow in designing a capture hood are:
1) Observe the operation through several cycles and interview workers and maintenance personnel about access
needs, work practices, materials handling, emergency
conditions, and maintenance.
2) Channel the airflow as much as possible by employing
flanges and placing the work on a horizontal surface.
Put a baffle to the rear and top if possible. Use side barriers only if the distance is great and the airflow is rel-
FIGURE 6-29. Overhead canopy hoods
6-28
Industrial Ventilation
7) When the hood is installed (and periodically thereafter), evaluate its performance both for ventilation
effectiveness and worker acceptance.
6.7
CHOOSING BETWEEN CAPTURE AND
ENCLOSING HOODS
When contaminants are toxic or are created at high velocities
and/or internal energy, then enclosing hoods are preferred.
If the contaminant source is considerably less hazardous
and manual access is required, the choice is less clear. Plug
flow enclosing hoods require greater care in design and
operation but are likely to be less vulnerable to poor work
practices. If the worker must move the capture hood frequently
for it to control the source, it is probably best to use a large
enclosing hood or a capture hood so that it need not be moved.
The main disadvantages of using enclosing hoods are initial
expense and the fact that they usually require more floor space.
Visual access and material handling issues may also suggest
use of a capture hood.
The main advantages of using capture hoods when
compared to enclosing hoods are that they: 1) require less
airflow if they are small and close to the source, 2) typically
can be used without modifying materials handling, 3) are less
expensive to purchase or build, and 4) require much simpler
selection, design, and installation.
Capture hoods can be extremely effective if the contaminant
is released: 1) with little or no initial velocity, 2) within the
hood’s effective range, and 3) at locations with relatively low
velocity (crossdrafts). The disadvantages of using capture
hoods compared to enclosing hoods are that their performance
typically can be strongly degraded more by: 1) seemingly
small changes in positioning of either the source or the hood,
2) crossdrafts and other competing air motions, and 3)
significant reductions in exhaust airflow. Because of their
greater control reliability, enclosures are always preferred over
capture hoods in situations where it is possible to install them.
This becomes critically important when dealing with
hazardous and toxic dust and vapors.
6.8
ERGONOMIC CONSIDERATIONS FOR DESIGN
OF HOODS
If workers must frequently reach into a hood or stand in it
to work, ergonomics and human factors should be employed
to make the hood as user-friendly as possible. For the
dimensions of the hood and work surfaces, flexibility in design
is a key ergonomic consideration since different workers with
varying physical characteristics may use the same workstation
over time. Hence, it is highly desirable in many cases to make
work heights and other critical dimensions adjustable.
Hoods, especially enclosing hoods, should allow clear sight
lines and sufficient light (without glare) for the work task.
Both reach-in and occupied hoods must be convenient and
comfortable for the worker to use. The width and height of the
hood should be large enough that the worker can safely handle
materials or equipment inside it. Usually, this will result in a
minimum width of at least 3 ft [0.9 m]. If the worker must lean
into the hood and lift relatively heavy objects the hood should
be wider. Occupied hoods should be at least 6 ft [approx. 1.8
m] wide to reduce claustrophobic reactions and allow room for
swinging the arms and bending the torso to the left and right.
If the user must spray or access the side of large objects
while within an enclosing hood, the hood width should be
greater than or equal to the width of the object plus 3 ft [0.9 m]
on each side.
If workers will enter a hood, the hood’s height must be
sufficient to allow headroom. A height of 7 ft [2.1 m] will
usually provide adequate headroom if the worker will not be
doing anything that requires moving the arms, materials or a
tool over the head.
Similarly, the heights of work tables, the floor of bench top
hoods, and other work surfaces should be set to accommodate
all workers and platforms used. If possible, working heights
should be adjustable by the worker. Convenient visual access
is also important. For example, the small enclosing hood
shown in Figure 6-30 should allow necessary sight-lines even
though it is relatively small. Like most other enclosing hoods,
the inside of the hood should be well lit and without glare.
Other considerations for hood design:
1) Provide a lean bar and foot rail where appropriate.
2) Have built in holders for tools and supplies (e.g., welding rods).
3) Suspend and counter balance heavy cables, tubing, etc.
that the worker must move around (e.g., electrical lines
for welders).
4) Counter balance movement arms for mobile hoods
(e.g., commercial welding hoods).
5) For enclosing hoods, use transparent plastic, tempered
glass or laminated safety glass for sides to allow visual
communication with nearby co-workers or to see items
that must be kept under surveillance (e.g., gauges and
indicators).
6) Move controls and indicators (e.g., hood static pressure
display) so they are close and visible while not interfering with the task.
7) Accommodate both left- and right-handed workers.
8) Avoid sharp or abrupt edges.
9) Provide noise protection from sources outside hood.
10) Where necessary, use safety controls such as Hands Off
buttons, dead man switches, etc.
11) Place outlets and controls for required utilities (compressed air, water, coolants, etc.) at convenient and safe
locations.
Hood Design
12) Consider ease of required cleaning or decontamination
tasks within and near the hood when selecting materials and drainage.
Design considerations for large hoods include:
1) Prevent heavy doors, sashes, work materials, etc., from
falling by using safety cables and counterweights.
2) Locate doors and sashes for easy access to enclosures
for both routine operations and maintenance.
3) Provide observation windows in doors to prevent collisions and to allow visual inspection of the inside.
4) If the inside of the hood would be hazardous during
operations, provide lockouts, interlocks, and warning
lights as needed.
Enclosing hoods sometimes also act as machine guards. In
those cases, additional regulations and guidelines may apply to
design.
Note: The intent of this section is to provide an overview of
ergonomic considerations for industrial ventilation design. If
specific ergonomic design details are needed, Universitybased Ergonomics Centers are resources, such as those at
Texas A&M University, North Carolina State University and
the University of Michigan.
6.9
6-29
WORK PRACTICES
Hoods used at or near workstations (i.e., bench top
enclosures and occupied hoods) should be designed and
operated with strong consideration of work practices. In some
cases work practices should be modified to accommodate the
hood but proper hood design includes consideration for work
practices and reliable continued safe use by the worker.
6.10
MATERIAL HANDLING IN AND NEAR HOOD
WORKSTATIONS
Moving products or materials into and out of the hood must
be convenient. Hood design, material handling, and work
practices should be considered as an integrated package.
Design techniques such as the chain slot (Figure 6-31) can
be used to accommodate both hood effectiveness and material
handling. Similarly, a sliding work table (Figure 6-32) or
turntable (Figure 6-33) can be used to solve both ventilation
and material handling issues.
Hoods mounted on articulating arms should be as
lightweight as possible and have full coverage of tasks being
ventilated.
Note that large objects in the hood can act as baffles and
stagnant zones near them. Proper design must consider these
effects and ensure contaminants are not recirculated back into
breathing zones.
The volume of gasses and dusts created in the process will
also affect hood design. In some cases they may overpower
some designs (such as downdraft booths) even when supply
air is provided above the process.
6.11
HOOD MAINTENANCE AND CLEANING
Access for maintenance of the hood and equipment within
should be considered in the initial hood design. This includes
provisions for lighting and utilities.
Collection of material should be convenient whenever
particulates (e.g., dusts or mist) may settle inside the enclosure
or the plenum of a slot plenum hood. For liquids, provide
inclined pathways to a drain (Figure 6-34). For dusts,
relatively steep slopes should be used where possible to
encourage settled material to slide to the bottom (Figure 6-35).
Ensure that materials that accumulate are not a safety or fire
hazard.
6.12
FIGURE 6-30. Small enclosing hood
HOODS AND PERSONNEL FANS
Personnel (or man-cooling) fans are usually located on the
floor or wall-mounted to provide personal comfort in warm
environments. They can move large volumes of air at high
velocities. Velocities of 2500 fpm [12.7 m/s] have been
recorded in front of personnel fans. That air movement can
overwhelm enclosing and capture hoods. Even if the flow is
perpendicular to the hood face, it is likely to reduce the
6-30
Industrial Ventilation
FIGURE 6-31. Chain slot
FIGURE 6-32. Roll out hood
effectiveness of the hood and promote the escape of
contaminant.
designs found earlier in this chapter.
Although personnel fans are likely to reduce the
effectiveness of both enclosing and capturing hoods, the hood
user might not be overexposed to airborne contaminants. The
fan may simply transport the contaminant away and cause it to
mix with the ambient air of the room.
Glove boxes should be used for high activity alpha or beta
emitters and highly toxic and biological materials. The air
locks used with the glove box should be exhausted if they open
directly to the room. For low activity radioactive laboratory
Replacing or removing personnel fans in areas where they
are established may be difficult to implement. In those cases,
special considerations in the hood design may be required to
keep exposures below acceptable values. It would also be
prudent to investigate other means to reduce heat stress (see
Chapter 10, Section 10.8).
6.13
VENTILATION OF RADIOACTIVE AND HIGH
TOXICITY PROCESSES
Ventilation of radioactive and high toxicity processes
requires knowledge of the hazards, use of extraordinarily
effective control methods, and adequate maintenance that
includes monitoring. Consult other resources, including the
published requirements of regulatory agencies for guidance.
Local exhaust hoods should be of the enclosing type with
the maximum containment possible. Where complete or
nearly complete enclosure is not possible, control velocities
from 50% to 100% higher than the minimum recommended
values in this Manual should be used. Laminar-flow supply air
should be introduced at low velocity and in a direction that
does not cause disruptive crossdrafts at the hood opening. This
is similar to operating room ventilation and clean room
FIGURE 6-33. Turntable
Hood Design
6-31
FIGURE 6-34. Dip tank hood that drains condensed fluid from plenums
work, a laboratory hood may be acceptable. For such hoods, a
minimum average face velocity of 80–100 fpm [0.41–0.51
m/s] is recommended (see Chapter 13, Section 13.35, VS-3501, VS-35-02, VS-35-04 and VS-35-20).
For preliminary design, air conditioning loads and other
requirements, consult the guidelines provided in Chapter 13,
Section 13.35 for hood airflow and replacement airflow. Other
regulatory standards should also be consulted. These values
may need to be revised as design conditions are firmed.
6.14
DETERMINING HOOD STATIC PRESSURE LOSS
Air flowing through a hood will cause pressure changes that
must be considered when connecting the hood to the system
duct. The sum of these changes is called hood static pressure
(SPh). The pressure losses in the hood are a function of the
speed of the air as it goes from “zero” to full duct velocity (i.e.,
FIGURE 6-35. Hopper bottom to ease removal of settled
materials
acceleration of the air) and the offering resistance. These
pressure losses are determined as a function of a loss factor
multiplied by the velocity pressure and include:(6.9,6.10)
1) The shape of the hood: The path the air takes as it
enters and moves through the hood and then into the
duct can be influenced by baffles, air turns through
plenums and boxes, and the raw edges of the hood
itself. Inlets can be designed to reduce these effects
(e.g., choosing bell mouth or tapered openings versus
plain openings). The hood shape itself can have gentle
turns and rounded paths to reduce these effects. A hood
entry loss factor (Fh), which is a function of the hood
shape, is used to describe the aerodynamic losses associated with moving air into duct (Figure 6-37 and
Sections 6.8 and 6.9). These aerodynamic losses occur
because air momentum within the hood carries the air away
FIGURE 6-36. Separation of flows at the duct inlet and
hood loss factors; values shown for round entries
6-32
Industrial Ventilation
from the walls of the hood and duct system. As long as
the shape of the hood is maintained (e.g., no cardboard
added to face or other physical alterations, etc.), this factor should remain constant. This pressure loss would be
determined by multiplying Fh by the duct velocity pressure (VPd).
In a compound hood, an additional loss factor, the slot
loss factor (Fs) is required to address the pressure loss
associated with moving air through the slot. The slot
velocity pressure (VPs) would be multiplied by Fs to
determine the pressure loss to move air through the slot
opening.
2) Acceleration of air: The air speed must increase from
zero (or some minimal) velocity outside the hood up to
duct velocities of over 1,000 fpm [5 m/s], and sometimes as high as 5,000–6,000 fpm [25.4–30.4 m/s].
This transfer of energy from static to kinetic is approximately equal to the pressure value of one velocity
pressure (VP) and must be included when calculating
the hood static pressure. The duct velocity pressure is
most commonly used to determine this loss. However,
in cases where the velocity of air moving through the
slot of a compound hood exceeds that of the air in the
duct following the hood, the slot velocity pressure
(VPs) would be used to determine this pressure loss.
The pressure loss associated with the conversion of
potential to kinetic energy resulting in the acceleration
of air is commonly called the acceleration or Bernoulli
loss (Fa) and equals 1.0VP.
3) Hood fittings: Fittings and/or filters associated with the
hood increase the pressure loss as air moves through
the hood and into the duct. Filters may be included at
the hood to remove large dry or liquid particles before
entering the duct system; they may also help with spark
arresting or keeping rags and other debris from getting
to the air-cleaning device. Not all hood systems include
fittings or filters. The filter loss can be separated from
the other hood components and read by a gauge mounted across the filter media (Figures 6-37a and 6-37b).
Filter manufacturers provide the filter loss as a ΔP
across the media (SPf) in "wg [Pa].
Taking into account the above information and
remembering that, by definition, the hood static pressure is a
negative value, one can estimate SPh by use of the following
equation:
SPh = -[(FhVPd + FsVPs + 1VP) + SPf]
where:
SPh = hood static pressure, "wg [Pa]
Fh = hood entry loss factor (dimensionless)
VPd = duct velocity pressure, "wg [Pa]
[6.9]
FIGURE 6-37a. Measurement location for SPfilter in typical
enclosing hood
Fs = slot entry loss factor (dimensionless)
VPs = slot velocity pressure, "wg [Pa]
VP = velocity pressure (when appropriate, use
the larger of VPd or VPs), "wg [Pa]
SPf = special fitting loss, "wg [Pa]
The value of SPf for a filter in a hood (Figures 6-37a and 637b) will vary. A minimum value will be realized when the
hood is new or recently cleaned, and a maximum value will be
realized when it should be replaced or cleaned. When
computing SPh for purposes of sizing ducts when there are
many hoods in the system and one or more has a filter, it is
advisable to use the middle of the range of values. If it is a onebranch system or a multiple branch system in which all filters
will be replaced or cleaned at once, then the maximum value
of the range should be used for fan selection.
To evaluate proper hood performance, one may monitor the
hood static pressure by placing a pressure gauge
approximately three duct diameters downstream of the hood’s
connection to the duct.
6.14.1 Static Pressure Losses for Simple Hoods. A
simple hood (i.e., not containing slots and a plenum) is
presented in Example Problem 6-2. If the hood face velocity
for a simple hood is less than 1,000 fpm [5.08 m/s], the slot
loss component of the hood static pressure (see Equation 6.10)
Hood Design
6-33
SPh = -[FhVPd + FaVPd + SPf]
SPh = -[(0.25)(0.56) + 0 + (1)(0.56)]
= -0.70 "wg
[SPh = -(0.25)(140) + 0 + (1)(140) = -175 Pa]
6.14.2 Pressure Loss in Compound Hoods. Example
Problem 6-3 illustrates how air flows through a double entry
loss (compound) hood. This is a double slot hood with a
plenum and a transition from the plenum to the duct. The
purpose of the plenum is to give uniform velocity across the
slot opening. Air enters the slot, in this case a sharp-edged
orifice, and loses energy due to the vena contracta at this point.
For this type of hood, losses occur at both the slot and the duct
entry (Figure 6-38).
FIGURE 6-37b. Measurement locations for SPfilter with filter
at entrance to hood and at the plenum face
EXAMPLE PROBLEM 6-3 (Compound Hood Loss)
Given: Compound hood taper entry angle = 45°
will be negligible. Therefore, simple hoods with face
velocities less than 1,000 fpm [5.08 m/s] and containing no
special fittings permit Equation 6.9 to be simplified:
SPh = -[(Fh + 1)VPd]
[6.10]
(Note: For the same type of hood consisting of a slot or
orifice but no plenum, Equation 6.9 becomes SPh = -[FhVPd +
FsVPs + FaVP]. Use the greater of VPd or VPs to determine the
acceleration loss.)
No hood filter (SPf = 0)
df = 1.0
Slot Velocity (Vs) = 2,000 fpm [10.16 m/s]
Duct Velocity (Vd) = 3,500 fpm [17.78 m/s] (Vd is greater than
Vs; therefore, apply the Bernoulli coefficient (Fa) to the duct
entry).
VPs = df (Vs/4,005)2 = (1.0)(2,000/4,005)2 = 0.25 "wg
[VPs = df (Vs/1.29)2 = (1.0)(10.16/1.29)2 = 62 Pa]
Fs, for slot (sharp edged orifice)
= 1.78 (from Chapter 9, Figure 9-a)
EXAMPLE PROBLEM 6-2 (Simple Hood Loss)
VPd = df (Vd/4,005)2 = (1.0)(3,500/4,005)2
= 0.76 "wg
Given: Simple hood, taper entry angle = 45°
No hood filter (SPf = 0); no slots
[VPd = df (Vd/1.29)2 = (1.0)(17.78/1.29)2 = 190 Pa]
= 0.25 as shown in Figure 6-36
Face Velocity (Vf) = Q/Af = 250 fpm [1.27 m/s]
Fh
Duct Velocity (Vd) = Q/Ad = 3,000 fpm [15.24 m/s]
SPh = -[FsVPs + FhVPd + SPf + FaVPd]
Fa/1.0
SPh = -[(1.78)(0.25) + (0.25)(0.76) + 0 + (1)(0.76)]
df = 1.0 (see Chapter 3)
VPd = df (Vd/4,005)2 = (1.0)(3,000/4,005)2 = 0.56 "wg
[VPd = (1.0)(15.24/1.29)2 = 140 Pa]
Fh = 0.25 as shown in Chapter 9, Figure 9-a
SPh = -1.40 "wg
[SPh = -[(1.78)(62) + (0.25)(190) + (1)(190)] = -348 Pa]
6-34
Industrial Ventilation
FIGURE 6-38. Compound losses in slot/plenum hood
6.14.3 Coefficient of Entry and System Evaluation. Ce,
also known as the Coefficient of Entry, is an important tool to
evaluate the operation of a system, primarily at hoods. Ce is
defined as the ratio between actual flow and the theoretical
flow possible under perfect conditions. For hood design, a
ratio of 1.0 would occur when Fh = 0 and all of the Hood Static
Pressure (SPh) is completely transferred to Velocity Pressure.
In practicality, this cannot exist but a value for Ce can be
determined either by calculation (if Fh is known) or by
measurement in the field. In effect, Ce is a measure of the
efficiency of the hood. Values for Ce can be any positive value
less than 1.0 and the hood is most efficient as the value
approaches 1.0.
Once the value of Ce is known, it can be used to provide
quick estimates of airflow (Q) by taking field measurements of
Static Pressure (SP) instead of performing normal traverses of
Velocity Pressure (VP) in the duct and complete subsequent
calculations. SP data taken can then be compared with a
baseline SP previously determined at startup.
Note: This technique is accurate only if there have been no
physical changes made to the hoods after measurements of SP
at startup (i.e., addition or relocation of baffles, slot
dimensions altered, or alterations to any hood dimension).
This includes temporary changes such as addition of
cardboard at openings, material buildup, etc.
6.14.4 Determination of Ce. Ce can either be calculated or
measured. Calculated values are not useful as a tool for system
evaluation and are more theoretical. This section will
concentrate on the use of a measured value and the quick
estimation of flow from that data.
The Ce can be determined, at startup of the system, by
measuring the SPh and the Duct Velocity Pressure (VPd). Since
the definition of Ce compares the ratio between actual and
theoretical flow, the square root of the ratio of these pressures
is used to make this calculation:
[6.11]
Note that this value will always be a positive value below
1.0. Once this hood efficiency is determined it can be used
when taking future measurements of SPh and inserting that
value in the Continuity Equation (Q = VA) and Chapter 3,
Equation [3.17b] plus inserting the VPd for a VP at any location:
Hood Design
6-35
[6.12] IP
[6.12] SI
From Equation 6.11 above:
and Equations above can be combined:
[6.13] IP
[6.13] SI
where:
Q = Airflow, acfm [m3/s]
A = Area, ft2 [m2]
SPh = Hood Static Pressure, "wg [Pa]
df = density factor (dimensionless)
Ce = Coefficient of Entry (dimensionless)
4,005, 1.29 (constants: IP, SI)
Coefficients of Entry are shown for a number of hood types
in Chapter 9, Figure 9a. These are calculated values and
should be considered estimates. Hood construction variations
and actual field conditions may alter hood design and
operating characteristics. Values for Ce should be determined
during system commissioning by actual measurements using
Equation 6.11. It must be noted that “Ce” flow determination
with hoods containing a hood filter is inappropriate as the filter
static pressure will continually change with operation.
EXAMPLE PROBLEM 6-4 Hood Flow Calculation (Use of
Ce to Calculate a new Q)
A ventilation technician at a lead-acid battery manufacturer must take quarterly airflow measurements in their dust
collection system per the OSHA Occupational Lead Standard
1910.1025 in order to protect the worker. At commissioning,
the system met the recommended minimum duct velocity
found in Chapter 5, Table 5-1. Measurements taken were SPh
(-2.2 "wg [-548 Pa]) and an average velocity that yielded a VPd
of 1.27 "wg [316 Pa]) in the 6” [0.160 m] diameter duct
connected to the hood; a Ce was calculated at 0.76, based on
an original flow rate of 882 acfm [0.42 am3/s]. Three months
later, an SPh was measured at -1.7 "wg [-423 Pa] and no
physical changes had been made to the hood. What is the
new airflow (Q)? Density Factor was 1.0 in all cases.
EXAMPLE PROBLEM 6-5 (Hood Flow Calculation [Use of
Ce to calculate Q])
Ce = 0.76 (Calculated during system start-up and operations
using Equation 6.11)
SPh = -1.15 "wg [-286 Pa] (measured at later date) 6" [0.160
m] diameter duct area = 0.1963 ft2 [0.020 m2]
df = 1.0
REFERENCES
6.1
Sanders, M.S.; McCormick, E.J.: Human Factors
Engineering, 7th Edition. McGraw-Hill Book
Company, New York (1993).
6.2
Caplan, K.J.; Knutson, G.W.: ASHRAE Trans.84(I),
511–521 (1978).
6.3
Guffey, S.E.; Barnea, N.: Effects of Face Velocity,
Flanges, and Mannikin Position on the Effectiveness
of a Benchtop Enclosing Hood in the Absence of
Cross-Drafts. Am. Ind. Hyg. Assoc. J. 55(2):132–139
(1994).
6.4
Brandt, A.D.: Industrial Health Engineering. John
Wiley and Sons, New York (1947).
6-36
Industrial Ventilation
Hood Design
6.5
Kane, J.M.: Design of Exhaust Systems. Health and
Ventilating 42:68 (November 1946).
6.6
Djamgowz, O.T.; Ghoneim, S.A.A.: Determining
Pickup Velocity of Mineral Dusts. Canadian Mining J.
(July 1974).
6.7
Silverman, L.: Velocity Characteristics of Narrow
Exhaust Slots. J. Ind. Hyg. Toxicology 24:276
(November 1942).
6.8
DallaValle, J.M.: Exhaust Hoods. Industrial Press,
New York (1946).
6-37
6.9
Brandt, A.; Steffy, R.: Energy Losses at Suction
Hoods. Heating, Piping & Air-Conditioning – Am.
Soc. Heat. Vent. Eng. J. Section, Sept: 105–119
(1946).
6.10
McLoone, H.E.; Guffey, S.E.; Curran, J.C.: Effects of
Shape, Size, and Air Velocity on Entry Loss Factors of
Suction Hoods. Am. Ind. Hyg. Assoc. J. 54(3):87–94
(1993).
6.11
Rossi, C.; Vargas, J.: CFD renderings in this Chapter
(2018).
Chapter 7
FANS
NOTE: Equations with notation followed by (IP) are designated for inch-pound system only; equations followed by (SI) are designated for
metric use only. If equation bears neither, then it applies to both systems.
CHAPTER SPECIFIC VOCABULARY . . . . . . . . . . . . . . . . . .7-3
FOREWORD . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7-4
7.1 INTRODUCTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7-4
7.2 FAN TYPES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7-4
7.2.1 Centrifugal Fans . . . . . . . . . . . . . . . . . . . . . . . . .7-4
7.2.2 Axial Fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7-5
7.2.3 Special Fan Types . . . . . . . . . . . . . . . . . . . . . .7-13
7.2.4 High Pressure Blowers and Vacuums . . . . . . .7-13
7.3 FAN SELECTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7-13
7.3.1 Fan Selection Criteria . . . . . . . . . . . . . . . . . . .7-13
7.3.2 Fan Selection Using Ratings Tables . . . . . . . .7-26
7.3.3 Fan Efficiency . . . . . . . . . . . . . . . . . . . . . . . . .7-27
7.3.4 Fan Selection at Non-Standard Density . . . . .7-27
7.4 FAN AND SYSTEM PERFORMANCE . . . . . . . . . . .7-32
7.4.1 Point of Operation . . . . . . . . . . . . . . . . . . . . . .7-33
7.4.2 Matching the Fan Curve and the System
Curve . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7-33
7.4.3 Affinity Laws for Fans and Systems . . . . . . . .7-35
7.4.4 Fan Affinity Laws Applied to Fan Curves . . . .7-35
7.4.5 Limitations on the Use of the Fan Affinity
Laws . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7-40
7.4.6
The System Affinity Laws Applied to
System Curves . . . . . . . . . . . . . . . . . . . . . . . . .7-40
7.4.7 Correlating System Static Pressure and Fan
Static Pressure to Power . . . . . . . . . . . . . . . . .7-44
7.5 FAN AND SYSTEM CONTROL . . . . . . . . . . . . . . . . .7-46
7.5.1 Flow Control Methods . . . . . . . . . . . . . . . . . . .7-46
7.5.2 Fans Operating in Series or Parallel . . . . . . . .7-51
7.6 FAN SYSTEM EFFECTS . . . . . . . . . . . . . . . . . . . . . . .7-54
7.6.1 Impact on System Performance . . . . . . . . . . . .7-54
7.6.2 System Effect Values . . . . . . . . . . . . . . . . . . . .7-54
7.6.3 Fan Inlet System Effects . . . . . . . . . . . . . . . . .7-55
7.6.4 Fan Outlet System Effects . . . . . . . . . . . . . . . .7-65
7.6.5 Calculating Fan System Effects . . . . . . . . . . . .7-65
7.7 FAN MOTORS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7-71
7.7.1 Motor Selection Criteria . . . . . . . . . . . . . . . . .7-71
7.7.2 Motor Installation . . . . . . . . . . . . . . . . . . . . . . .7-73
REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7-73
ACKNOWLEDGMENTS . . . . . . . . . . . . . . . . . . . . . . . . . . . .7-73
____________________________________________________________
Figure 7-1
Figure 7-2
Figure 7-3
Figure 7-4
Figure 7-5
Figure 7-6
Figure 7-7
Figure 7-8
Figure 7-9
Centrifugal Fans . . . . . . . . . . . . . . . . . . . . . . . . .7-5
Axial and Special Types of Fan Designs . . . . . .7-7
Centrifugal Fan, Exploded View . . . . . . . . . . . .7-9
Propeller Fan, Exploded View . . . . . . . . . . . . .7-10
Tubeaxial Fan, Exploded View . . . . . . . . . . . .7-11
Vaneaxial Fan, Exploded View . . . . . . . . . . . .7-12
Tubular Centrifugal Fan, Exploded View . . . .7-14
Air Ejectors . . . . . . . . . . . . . . . . . . . . . . . . . . .7-15
AMCA 99-16 Classification for Spark
Resistant Construction . . . . . . . . . . . . . . . . . . .7-17
Figure 7-10 Rotation and Discharge Configurations of
Centrifugal Fans . . . . . . . . . . . . . . . . . . . . . . . .7-18
Figure 7-11 Centrifugal Fan Drive Arrangements . . . . . . .7-19
Figure 7-12 Centrifugal Fan Drive Arrangements,
Continued . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7-20
Figure 7-13 Axial Fan Drive Arrangements . . . . . . . . . . . .7-21
Figure 7-14 Tubeaxial and Tubular Centrifugal Fan Drive
Arrangements . . . . . . . . . . . . . . . . . . . . . . . . . .7-22
Figure 7-15IP Estimated Belt Drive Loss . . . . . . . . . . . . . . . .7-23
Figure 7-15SI Estimated Belt Drive Loss . . . . . . . . . . . . . . . .7-24
Figure 7-16
Figure 7-17
Figure 7-18
Figure 7-19
Figure 7-20
Figure 7-21
Figure 7-22
Figure 7-23
Figure 7-24
Figure 7-25
Figure 7-26
Figure 7-27
Figure 7-28
Figure 7-29
Figure 7-30
Figure 7-31
Figure 7-32
AMCA Fan Classes, Airfoil and Backward
Inclined Single Width . . . . . . . . . . . . . . . . . . .7-29
Curing Process . . . . . . . . . . . . . . . . . . . . . . . . .7-31
Typical Fan Performance Curve . . . . . . . . . . .7-34
System Curves . . . . . . . . . . . . . . . . . . . . . . . . .7-35
System Curves . . . . . . . . . . . . . . . . . . . . . . . . .7-36
Actual vs Desired Point of Operation . . . . . . .7-37
AMCA Air Performance with Certified
Ratings Tolerance . . . . . . . . . . . . . . . . . . . . . . .7-38
Fan Selection at Standard Conditions . . . . . . .7-39
Effect of 10% Increase in Fan Speed . . . . . . .7-40
Effect of 50% Decrease in Gas Density . . . . .7-41
Homologous Fan Performance Curves . . . . . .7-42
Typical Backward-Inclined Fan Curves
with Volume Controls . . . . . . . . . . . . . . . . . . .7-47
Constant Torque . . . . . . . . . . . . . . . . . . . . . . . .7-48
Variable Torque . . . . . . . . . . . . . . . . . . . . . . . .7-49
Controls and Power Comparison . . . . . . . . . . .7-49
Flow Control – Constant Speed . . . . . . . . . . . .7-50
Flow Control – Variable Speed . . . . . . . . . . . .7-50
7-2
Industrial Ventilation
Figure 7-33 Direct Drive Fan With VFD Control . . . . . . . .7-51
Figure 7-34 Fans Series Operation . . . . . . . . . . . . . . . . . . .7-52
Figure 7-35 Fans Parallel Operation . . . . . . . . . . . . . . . . . .7-53
Figure 7-36 Fan System Effect (FSF) . . . . . . . . . . . . . . . . .7-54
Figure 7-37IP System Effect Losses . . . . . . . . . . . . . . . . . . . .7-56
Figure 7-37SI System Effect Losses . . . . . . . . . . . . . . . . . . . .7-57
Figure 7-38 System Effect Factors . . . . . . . . . . . . . . . . . . .7-58
Figure 7-39 Non-Uniform Inlet Flows for
Centrifugal Fans . . . . . . . . . . . . . . . . . . . . . . . .7-59
Figure 7-40 Non-Uniform Fan Inlet Corrections . . . . . . . .7-60
Figure 7-41 Round Inlet Elbows for Centrifugal Fans . . . .7-61
Figure 7-42 Rectangular Inlet Elbows for Centrifugal
Fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7-62
Figure 7-43
Figure 7-44
Figure 7-45
Figure 7-46
Figure 7-47
Figure 7-48
Figure 7-49
Figure 7-50
Figure 7-51
Inlet Elbows for Axial Fans . . . . . . . . . . . . . . .7-63
Inlet Vane Dampers for Centrifugal Fans . . . .7-64
Obstructed Fan Inlets . . . . . . . . . . . . . . . . . . . .7-66
Fan Outlet Ducts for Centrifugal Fans . . . . . .7-67
Fan Outlet Ducts for Axial Fans . . . . . . . . . . .7-68
Fan Outlet Elbows for Centrifugal Fans . . . . .7-69
Fan Outlet Elbows for Axial Fans . . . . . . . . . .7-70
Fan Inlet Elbow (Example Problem 7-9) . . . .7-71
Motor Locations for Belt Driven
Centrifugal Fans . . . . . . . . . . . . . . . . . . . . . . . .7-73
____________________________________________________________
Table 7-1(lP) Example of Multi-Rating Table . . . . . . . . . . . .7-28
Table 7-1(SI) Example of Multi-Rating Table . . . . . . . . . . . .7-28
Fans
7-3
Chapter Specific Vocabulary
AMCA: Air Movement and Control Association, Arlington
Heights, IL
Constant Torque Load: The energy demand on a motor by a
constant torque machine such as a rotary valve or screw
conveyor. When operating on a variable frequency drive, the
drive is configured for a linear volts to hertz ratio where from
0–60 Hz the motor output power varies linearly with the change
in motor speed while torque remains constant.
Equivalent Pressure (Peqv): Pressure at actual conditions
corrected to standard conditions, 0.075 lb/ft3 air stream density
[1.204 kg/m3].
Fan: A constant flow device used to displace air at low to
medium pressures based on the air stream density at the fan
inlet. Fans having one or more wheels (impellers) in parallel on
a common shaft are called single stage fans, while fans having
more than one wheel (impeller) in series on a common shaft and
in a common housing are called multi-stage fans.
Fan, Tubular Centrifugal: A fan utilizing an in-line, axial type
housing and a centrifugal wheel in which the air enters the
wheel axially, exits the wheel at a 90° angle and is then turned
back in the axial direction by internal turning vanes downstream
of the wheel.
Flow Rate (Q): The airflow in volume per unit of time used for
system design and fan selection stated in actual cubic feet per
minute (acfm), standard cubic feet per minute (scfm) or dry
standard cubic feet per minute (dscfm) [actual cubic meters per
second, acms; normal cubic meters per second, ncms; dry
normal cubic meters per second dncms].
Hybrid Flow System: A system with one or more of its
components operating in non-turbulent flow conditions.
Noise, Aerodynamic: Fan noise created by aerodynamic
effects such as air turbulence and fan blade pass frequency.
Fan Affinity Laws: Relationships between fan flow rate,
speed, pressure and power used to calculate and predict
additional points of operation from one fan curve to one or more
additional fan curves.
Noise Directivity and Distance: Noise directivity is the
number of radiating surfaces reflecting fan sound at a specified
distance. Directivity is normally referred to as Q1, Q2, Q4 or Q8
and distance is in either feet or meters.
Fan, Axial: A fan in which the air travels parallel to and along
the axis of the fan shaft. Axial fans may be housed or unhoused.
Noise, Mechanical: Fan noise created by mechanical effects
such as the motor, drive components and bearings.
Fan, Centrifugal: A fan in which the air enters the fan wheel
axially and exits the wheel at a 90° angle. Centrifugal fans may
be housed or un-housed.
Fan Drive Arrangement: The location of the motor with
respect to the fan when viewed from the drive side of the fan.
Fan Noise: A type of unwanted sound produced by the fan or
its components that might require corrective action to bring the
fan sound to an acceptable level.
Fan Rotation: The direction of the wheel rotation, CW or
CCW, when viewed from the drive side of the fan.
Fan Sound Power (Lw): Expressed in decibels (dB), the
sound power energy produced by the fan and used for fan
sound ratings. Fan sound power levels are the basis of AMCA
certification for fan sound performance.
Fan Sound Pressure (Lp): Fan sound power converted to an
estimated sound pressure value using the A or C scale. Fan
sound pressures are not AMCA certified.
Fan Static Efficiency (FSE): The fan energy efficiency based
on static pressure.
Fan Static Pressure (FSP): The static pressure used for fan
ratings and defined by AMCA as the fan total pressure minus the
velocity pressure at the fan outlet.
Fan System Effect: The estimated loss in fan performance due
to non-uniform flow conditions at the fan inlet or outlet.
Fan Total Efficiency (FTE): The fan energy efficiency based
on total pressure. Sometimes referred to as Mechanical
Efficiency.
Fan Total Pressure (FTP): The increase in total pressure
across the fan and defined by AMCA as the total pressure at the
fan outlet minus the total pressure at the fan inlet.
Point of Operation: The specified flow rate and pressure that
either a fan or a system is intended to perform.
Power (PWR): The energy demand (load) expressed in
horsepower, watts or kilowatts. When used for fans, it is referred
to as fan shaft power (FAN-PWR), when used for drives, it is
referred to as drive power (DRV-PWR) and when used for
motors it can be referred to as either motor input or motor output
power (MTR-PWRinput or MTR-PWRoutput). The motor output
power must be equal to or greater than the sum of the fan and
drive power; the motor input power must be equal to or greater
than the motor output power.
System Affinity Laws: Relationships between system flow
rate, pressure and power used to calculate and predict
additional points of operation from one system curve to one or
more additional system curves.
System Effect Factor (SEF): A loss factor specific to a
system component that is multiplied by the component velocity
pressure to calculate the fan system effect loss.
System Effect Loss (SEL): The static pressure loss ("wg or
Pa) for a system component causing a fan system effect.
Turbulent Flow System: A system that operates wholly in
turbulent flow conditions.
Variable Torque Load: The energy demand on a motor by a
variable torque machine such as a fan or pump. When operating
on a variable frequency drive, the drive is configured for a
squared volts to hertz ratio where from 0–60 Hz the motor
output power and torque varies with the square of the change in
motor speed.
7-4
Industrial Ventilation
FOREWORD
The Department of Energy (DOE) is formulating a new
Regulation on Energy Conservation Standards for Commercial and Industrial Fans and Blowers (10 CFR Part 431) governing fan energy efficiency. If passed, this regulation will
have a noticeable impact on how fans are rated and selected.
Due to its impact, the regulation is expected to have a five (5)
year implementation period for industry compliance. This proposed regulation is not discussed in this edition of the Manual
but if passed, will be addressed in a later edition.
7.1
INTRODUCTION
Fans utilized for industrial ventilation systems are typically
single stage, constant flow devices used to displace air at low
to medium pressures up to 140 "wg or 5 psig [34.9 kPa] based
on the air stream density at the fan inlet. Fans having one or
more wheels (impellers) in parallel on a common shaft are
called single stage fans, while fans having more than one
wheel (impeller) in series on a common shaft and in a common
housing are called multi-stage fans. Because the wheels are in
series, multi-stage fans are capable of generating higher pressures than single stage fans, and are often a viable selection for
higher pressure systems (> 140 "wg or 5 psig [34.9 kPa].
When fans are used to induce, or exhaust air through a system
they are often referred to as induced draft fans or exhausters.
When fans are used to push air through a system, they are
often referred to as forced draft fans or blowers. An example
of an induced draft fan is an installation on the clean side of a
dust collector exhausting air to atmosphere, while examples of
a forced draft fan are supplying fresh air into a burner system
or supplying fresh or conditioned air into a workplace. When
the fan is installed within a system with inlet and outlet ducting, it can be referred to as either an exhauster or blower
depending on the primary purpose of the system design.
Compressors typically work at medium to high pressures
above 5 psig or 140 "wg [34.9 kPa]) for positive pressure
applications or below -140 "wg or -10.3 "Hg [-34.9 kPa]) for
negative pressure applications. In most cases, the designer
loses efficiency when the selection is based solely on pressure.
Fans can be up to 90% efficient at low pressures, while some
compressors delivering low flow and high pressure may only
reach an efficiency of 10–35%.
In this Manual, we refer to the air moving device used in an
industrial ventilation system as a single stage fan, though in
some specialty applications other types of air movers such as
regenerative blower, multi-stage fan, rotary lobe vacuum or
pressure blowers or compressors might be used instead of a
fan.
The fan is an important component of the ventilation system
as it supplies both the energy to move the air through the system (velocity pressure) and the energy to overcome resistance
to flow (static pressure). The energy to move the air through
the system is based on the specified velocity, and the energy to
overcome resistance to flow is from system components such
as hoods, ducting, air cleaning device, etc.
The most widely used types of fans in industrial ventilation
systems are centrifugal fans, in which air exits the fan 90
degrees from entry, and axial fans, in which air exits the wheel
along the axis of the fan shaft.
This chapter describes the common types of centrifugal and
axial fans used for industrial ventilation, fan selection procedures, fan and system performance, fan system effects and
motors.
7.2
FAN TYPES
Fans are categorized into two types: centrifugal and axial
flow. Information on wheel and housing design, fan curves,
performance characteristics and typical applications is provided in Figures 7-1 and 7-2. The right side of these figures provides an explanation of the fan curve nomenclature.
7.2.1 Centrifugal Fans (Figure 7-3). A Centrifugal Fan
consists of a wheel or impeller mounted on a shaft, normally
rotating in a scroll shaped housing. The air enters the wheel
axially and exits at a 90 degree angle. The rotation of the wheel
imparts kinetic energy to the air between or along the blades.
This kinetic energy is converted to static pressure as the air
slows when exiting the wheel. Centrifugal fans have three
basic impeller designs: forward curved, radial and backward
inclined.
Forward Curved Wheels (FC or squirrel cage) use short,
cupped blades curved into the direction of rotation. These fans
are used in applications for low to moderate static pressures up
to about 5 "wg [1.25 kPa] such as in low pressure heating and
cooling equipment. The fan curve has a characteristic dip to
the left of the peak pressure followed by a gradual decline at
higher flow rates. Care must be taken to accurately calculate
the fan static pressure and select the fan to operate to the right
of the peak pressure to avoid unstable operation or pulsation.
Forward curved wheels have a horsepower curve that rises
with increasing flow rates and are often called overloading
type fans. This type of fan is not recommended for wet or dry
contaminant laden air streams that could bind to the blades
causing imbalance and reduced performance.
Radial Wheels (R) have blades that project straight or radially from the hub and are characterized by their simple and
rugged construction. The housings are designed with inlets
and outlets producing high velocities to keep contaminants
entrained in the air stream until exiting the fan. Since the air
moves across both sides of the blades, radial wheels tend to be
self-cleaning and can be constructed to withstand erosion and
impact damage from airborne contaminants. The performance
curve for a radial fan is often stable with a minimal dip and a
long, steep volume-to-pressure slope. The major disadvantage
of the radial blade is its lower efficiency and overloading
horsepower characteristic when compared to the backward
inclined wheel.
Fans
7-5
FIGURE 7-1. Centrifugal fans; impeller and housing designs
Backward Inclined (BI) Wheels have blades that are
inclined at an angle with the direction of rotation. Due to its
higher efficiency, backward inclined fans run at higher rotating
speeds than forward curved or radial fans for the same flow
rate and pressure. Being more efficient, backward inclined
fans normally operate at a lower noise level. The shape of the
horsepower curve is non-overloading, in that the maximum
horsepower peaks at the point of optimum efficiency and
declines at higher and lower flow rates. For this reason, these
fans are often referred to as “non-overloading” fans. This feature makes these fans a safe choice for motor selection when
the system static pressure fluctuates or when the design calculations are not well defined. While some blade shapes can handle light concentrations of contaminants or moisture, backward inclined fans are designed for clean air applications such
as on the clean side of high efficiency air pollution control
equipment.
Backward inclined wheels have three basic blade shapes –
single thickness-backward inclined, single thickness-back-
ward curved, and airfoil. With a heavier, single thickness
blade, the single thickness backward inclined and backward
curved blades are more durable than the airfoil and operate at
slightly lower speeds and efficiency.
The airfoil blade is shaped like the cross-section of an airplane wing, normally has a hollow core, and operates at higher
speeds and efficiency than single thickness blade wheels.
Typically hollow, the airfoil blade is subject to the effects of
abrasion, erosion and moisture and is limited to clean, dry air
streams. Because of the airfoil shape of the blades, materials of
construction are limited due to the forming stresses during
shaping. If an airfoil wheel is used in a humid air stream, consult the fan manufacturer regarding the use of “weep” holes in
each of the blades to relieve possible condensation on the
inside of the blades.
7.2.2 Axial Fans (Figures 7-4, 7-5 and 7-6). Axials are fans
in which the air travels through the fan along the axis of the fan
shaft. Axial Fans are generally best suited for handling medium to high volumes of relatively clean air at low static pres-
7-6
Industrial Ventilation
FIGURE 7-1 (cont.). Centrifugal fans; performance curves and characteristics. (*These performance curves reflect the general characteristics of various fans as commonly employed. They are not intended to provide complete selection criteria for
application purpose, since other parameters such as diameter and speed are not defined.)
sures and temperatures. Axials have an advantage in that they
are compact and move air in a straight line. They have limited
application when the air stream is dusty, corrosive or explosive, as the bearings and drive components may be partially
exposed to the air stream and the wheels are not suitable for
contaminant laden air streams. Three common types of axial
fans are: Propeller, Tubeaxial, and Vaneaxial, with the
Vaneaxial being the most efficient design.
Propeller Fans, also known as panel fans, are not ducted and
are used in applications moving large volumes of air against
low static pressures. Flow rates are very sensitive to added
resistance and an increase of 0.5 to 1 "wg [125 to 250 Pa] can
cause a marked reduction in flow. If the propellers have an airfoil design the fan can operate in the pressure range of 1 to 2
"wg [249 to 498 Pa]. Propeller fans are best suited for use as a
non-ducted roof or wall fan for circulating, exhausting or sup-
plying air for large spaces.
Tubeaxial Fans use airfoil or propeller type blades mounted
in a cylindrical housing. The spacing between the blades and
the housing impacts efficiency while the hub to blade diameter
ratio determines the pressure capacity. Tubeaxial fans move
low to high flow rates against moderate pressures up to about
3 "wg [747 Pa]. The tubeaxial fan can be ducted, which broadens its scope of application for use in low pressure, local
exhaust ventilation systems.
Vaneaxial Fans are tubeaxial fans with the addition of flow
straightening vanes inside the housing and just after the wheel.
This requires a longer housing but also provides higher efficiencies and pressure capacities up to 8 "wg [2 kPa]. Some
vaneaxial fans can be constructed to include manual or automatically controlled variable pitch blades to increase fan efficiency and capacity. Some designs can be modified for high
Fans
FIGURE 7-2. Axial and special types of fan designs; impeller and housing designs.
7-7
7-8
Industrial Ventilation
FIGURE 7-2 (cont.). Axial and special types of fan designs; impeller and housing designs.
Fans
7-9
7-10
Industrial Ventilation
Fans
7-11
7-12
Industrial Ventilation
Fans
temperature air streams and higher pressures.
7.2.3 Special Fan Types (Figures 7-7 and 7-8). Tubular
Centrifugal Fans use backward inclined wheels inside a cylindrical housing to create an in-line centrifugal fan allowing an
in-line duct installation and operation at higher capacities and
lower noise than typical axial fans (Figure 7-7). The air enters
the wheel axially, turns at a 90 degree angle, and then turns
again axially through the housing via straightening vanes. The
flow-pressure-power characteristics of this fan are similar to a
scroll type centrifugal fan of the same blade type except with
lower efficiencies. Space requirements are similar to vaneaxial
fans.
Mixed Flow Fans use wheels combining the characteristics
of axial and centrifugal wheels (Figure 7-7). The airflow
makes two 45 degree turns as it passes through the wheel. This
produces higher flow rates and pressures than the axial fan and
higher efficiencies than the tubular centrifugal fan. Mixed flow
fans are used for supply or return air or general ventilation
applications requiring higher flow rates at moderate pressures.
The fan’s main advantages are its high operating efficiencies
and lower noise levels.
Ejectors are sometimes used when it is not desirable to have
either contaminated or high temperature air or degradable or
abrasive materials pass directly through the fan. Ejectors are
used for air streams with corrosive, flammable, explosive, hot,
or sticky materials that might damage a fan or be damaged by
the fan, present a dangerous operating condition, or degrade
fan performance. The contaminated air is induced into the
ejector by the jet action of high pressure primary air, which is
usually from a secondary fan. A typical design is shown in
Figure 7-8. Ejectors have low efficiencies, often between
15–20%. Ejectors are also used in pneumatic conveying systems.
7.2.4 High Pressure Blowers and Vacuums. Regenerative
Blowers are sometimes used in applications requiring flow
rates from about 50 to 700 acfm [0.33 acms] and pressures up
to about 3.9 psig or 7.6 "Hg [26.9 kPa or 25.7 kPa].
Regenerative blowers are single stage devices and utilize a
compression space between the blade tips and the housing that
allows a portion of the air to migrate from the blade tip back
into a succeeding blade for reacceleration and increased pressure. For this reason, while they are very effective in generating high pressure or vacuum capacities, their efficiencies are
quite low, normally between 25–35%.
Multi-Stage Fans are often used for applications requiring
stable flow rates up to 40,000 scfm [18.9 scms] with pressures
ranging from 3 to 24 psig [20.7 kPa to 165.5 kPa] or 2 to 18
"Hg [6.8 kPa to 61.0 kPa]. Multi-stage fans utilize two or more
wheels installed in series (stages) on a common shaft inside a
common housing to generate medium to high pressures at stable flow rates without a significant loss in efficiency. The number or stages (wheels) can range from 2 to 11, and various
wheel types (backward inclined, radial, etc.) are often com-
7-13
bined to achieve the maximum efficiency and performance.
For this reason, multi-stage fans often achieve static efficiencies of 76 to 80%. Note that the efficiency Equations 7.4 and
7.5 in Section 7.3.3 are for single stage fans and are not directly applicable to high pressure blowers and vacuums. This is
because since efficiency is a function of heat loss, for single
stage fans the heat loss (e.g., compression, friction) is largely
insignificant compared to the change in pressure, so single
stage fan efficiency is calculated on the pressure basis. But at
higher pressures, heat loss becomes significant and efficiency
is calculated using heat (change in temperature) as the basis
instead of pressure.
7.3
FAN SELECTION
This section covers fan selection criteria and provides
guidelines for selecting a fan to meet the flow rate and static
pressure needs of the industrial ventilation system.
7.3.1 Fan Selection Criteria.
CAPACITY
Flow Rate (Q): Calculated by the system designer based on
the system airflow requirements and expressed as actual cubic
feet per minute (acfm) [acms] at the fan inlet.
Pressure Requirements: Calculated by the system designer
based on the system pressure requirements and expressed as
Fan Static Pressure (FSP) or Fan Total Pressure (FTP) at
actual conditions ("wg) [kPa, Pa], or as Equivalent Fan Static
or Total Pressure (FSPeqv, FTPeqv) corrected to standard conditions (0.075 lbm/ft3) [1.204 kg/m3] at the fan inlet.
AIR STREAM CONDITIONS
Material Handling: The type of fan selected is influenced
by the concentration and characteristics of the contaminant in
the air stream. Radial fans are the most common design for
material handling applications; however, backward inclined
fans can be used for low concentrations of dry, non-sticky and
non-abrasive dust, smoke and fumes. An axial fan can also
handle low concentrations of smoke and moisture, but is commonly used for those applications where large air volumes are
handled at lower static pressure requirements.
Toxic Gases: When the gas stream contains toxic gases the
fan construction should be such that there is no leakage into or
out of the fan. Consult the fan manufacturer to determine the
degree of airtight construction available for a specific fan
selection.
Explosive or Flammable Material: This may require spark
resistant fan construction and/or an explosion proof motor,
depending on the motor location or the area hazard classification. While motors can be classified by the National Electrical
Code (NEC) as explosion proof, fans are classified as spark
resistant according to Air Movement and Control Association
(AMCA) Standard 99 and National Fire Protection
Association (NFPA) regulations (Figure 7-9). When convey-
7-14
Industrial Ventilation
Fans
7-15
7-16
Industrial Ventilation
ing explosive or flammable materials, it is important to recognize the potential for ignition in the gas stream. This may be
from airborne material striking the fan wheel, the wheel slipping on the shaft, undissipated static electricity, etc. To minimize these dangers the fan may need special construction such
as a more secure attachment of the wheel to the shaft, bearing
stop blocks, the use of buffer plates, or spark resistant alloy
construction. Because no single type of construction fits all
applications, it is imperative that the user be aware of all dangers and specifies the type of construction and degree of protection required.
Note: For many years, aluminum alloy wheels have been
specified to minimize sparking against the wheel coming into
contact with other steel parts. This is still accepted, but tests by
the U.S. Bureau of Mines(7.1) and others have shown that the
impact of aluminum with rusty steel creates a “Thermite” reaction and possible ignition hazards. Special care must be taken
when aluminum alloys are used in the presence of steel.
Additionally, AMCA Standards for Spark Resistant Construction prohibits the use of any alloy having an iron content of
more than 5%. Hardware, however, such as set screws or keys,
may have an iron content greater than 5% provided they are
recessed and relatively inaccessible.
Corrosive Applications: This may require a protective coating or special materials of construction (stainless steel, fiberglass, etc.) to minimize corrosion. In applications where the
gas can condense, the fan housing can be insulated to minimize heat transfer.
Air Stream Temperatures: Minimum and maximum operating temperatures and rate of temperature change affect the
stresses and thermal growth of various parts of the fan. The
selection of correct materials of construction, arrangement,
and bearing types must take into account the temperature of
the gas stream. Most heat fans are designed for a maximum
temperature rise of 15 to 20 F [-9.4 to -6.7 C] per minute.
used to allow fan maintenance without disturbing the ducting,
and a fan inlet box can be used in lieu of an elbow at the fan
inlet when straight ducting is not possible.
FAN ORIENTATION
Fan Rotation (CW or CCW) is viewed from the motor drive
side of centrifugal fans and the fan outlet (discharge) end of
axial fans. The fan discharge position for centrifugal fans is
viewed from the motor drive side of the fan. Fan rotation and
discharge positions for centrifugal and axial fans are defined
by AMCA and ISO standards and are shown in Figures 7-10
and 7-14.
DRIVE ARRANGEMENTS
All fans must have some type of prime mover. This is usually an alternating current (AC) electric motor. In many cases,
the motor is furnished with the fan and factory mounted by the
manufacturer. In other cases, the motor is supplied separate
from the fan for site alignment and installation. Standard Fan
Drive Arrangements established by AMCA are shown in
Figures 7-11, 7-12, 7-13 and 7-14.
Direct Drive Fans couple the fan direct to the motor shaft
either by installing the fan wheel directly onto the motor shaft
or by coupling the fan shaft directly to the motor shaft. In both
cases, the fan operates at the motor speed. A direct drive fan
eliminates the need for a belt drive and has several advantages,
including higher efficiency (no drive losses), reduced noise
(no belt noise) and lower maintenance (no belts to adjust).
Except when using a variable speed drive controller, fan
speeds are limited to the available motor speeds, which are
typically 900 rpm, 1200 rpm, 1800 rpm or 3600 rpm at 60 Hz
[750, 1000, 1500 or 3000 at 50 Hz]. Fan capacity is varied by
either constructing the fan with a non-standard wheel geometry (width, diameter) or varying the motor speed via a variable
frequency drive controller, or both.
The best fan selection does not always fit in the space available. The fan type (centrifugal or axial) and speed chosen
determines the fan size. The fan size defines the space and
foundation requirements and the cost of the fan installation. In
other cases, important factors like fan weight and sound may
dictate whether the fan should be installed indoors, outdoors,
mounted on a concrete slab or placed remote from system
operations.
Belt Drive Fans use belts and sheaves to transfer power
from the motor shaft to the fan shaft. Belt drives allow the fan
speed to be changed with adjustment of the sheave diameters
or by changing the drive ratio. The drive ratio is the motor rpm
divided by the fan rpm. This may be important in some applications to provide for changes in system capacity or pressure
requirements when there are changes in the process, hood or
duct design, equipment location, or air cleaning equipment.
V-belt drives have drive losses that can be estimated from
Figure 7-15 using the drive loss factor, Fdrv.
The ability of personnel to access and service the fan is also
important and must be considered during the project design
phase. Space must be provided for the fan installation to allow
access for service and maintenance. The fan inlet and outlet
must be properly oriented and a sufficient length of straight
duct at the fan inlet and outlet should be provided to avoid fan
system effects and provide maintenance access to the fan
wheel and shaft (Section 7.6). A split housing design can be
Drive System Tolerances. For direct drive fans, fan selection is normally made using industry standard synchronous
motor speeds based on motor type and speed. For instance, a
direct drive fan selected using a 3600 rpm motor may actually
operate anywhere between 3480 and 3555 rpm depending on
the motor type, horsepower and manufacturer. If the fan performance is critical, the designer should consult the fan or
motor manufacturer for the actual speed of the selected motor
SPACE LIMITS
Fans
FIGURE 7-9. AMCA 99-16 Classifications for Spark Resistant Construction
7-17
7-18
Industrial Ventilation
Fans
7-19
7-20
Industrial Ventilation
Fans
7-21
7-22
Industrial Ventilation
Fans
7-23
7-24
Industrial Ventilation
Fans
when making the fan selection. For belt drive fans, although
fan selection software and tables allow fan selection at any
whole number rpm, in practice the fan manufacturer is limited
by the range of commercially available drive sets. For
instance, if the fan selection is at 2063 rpm and the closest
commercially available drive set is 2090 rpm, the fan manufacturer would normally supply the fan at 2090 rpm.
FLOW CONTROL
In most systems, a single fan is selected for the operating
condition. Some fans may be required to operate over a wide
range of flow rate or pressure requirements due to system variables. For example, a system may require the fan to provide a
constant flow rate based on a variable process condition or
changes to the system static pressure such as when the pressure drop across the baghouse filters varies. The fan may also
need to operate over a wide range of fan curves when new
hoods and ductwork are added or when correcting system
design deficiencies. Control methods include outlet dampers,
inlet box dampers, inlet vane dampers, variable pitch blades
(axials only) and variable speed drives (Section 7.5).
Sometimes it may be necessary to install two or more fans
in a system to provide a higher pressure or flow rate than can
be achieved with a single fan. Placing fans in series or parallel
may offer an advantage when it is necessary to boost static
pressure or volumetric flow rate. When fans are installed in
parallel, all fans are selected for the same static pressure but
may have the same or differing flow rates. In this case, the system flow rate is additive of each of the fan flow rates but the
static pressures are not. When fans are installed in series, each
fan is selected for the same flow rate (corrected for inlet densities) but the fan pressures may vary from fan to fan. In this
case, the static pressure is additive of each fan but the flow
rates are not. See Section 7.5 for fans operating in series and
parallel.
FAN SOUND
Aerodynamic and mechanical noise effects are the principal
sources of fan noise. The term fan noise describes a type of
unwanted sound produced by the fan or its components that
may require corrective action to bring the fan sound to an
acceptable level.
Aerodynamic Noise effects are from air turbulence and
blade pass frequency. Air turbulence is the most common
source of aerodynamic fan noise and can be transmitted
through system ducting upstream and downstream of the fan
and to external areas adjacent to the duct system in the form of
breakout noise. Sources of air turbulence include a) resistance
to flow is too high, resulting in insufficent flow into the wheel,
b) flow separation at the fan blades due to unstable or insufficient flow and c) sudden changes in the flow profile as air
approaches and moves through the fan. Blade pass frequency
is a pure tone produced as the fan wheel rotates past the fan
7-25
outlet in centrifugal fans and past the straightening vanes in
axial fans. This sound is similar to a moving object passing a
stationary object (such as an automobile passing a person
standing still) and is calculated by multiplying the number of
fan blades by the rotating speed of the fan (rpm). If this frequency matches or is close to the natural frequency of the connecting structure or ducting, it can excite the structure or ducting resonance and increase the sound level and, in some cases,
create destructive forces to the fan, structure or ducting.
Proper fan selection also has a siginificant impact on airborne sound levels. A common perception is that the slower the
fan speed, the lower the sound level, and while this is sometimes true, the lowest sound level is achieved by selecting and
operating the fan at the highest possible total efficiency.
Mechanical Noise effects are from fan components such as
the motor, drive and bearings that can transfer sound mechanically to the fan, structure or duct system. Since sound data
provided by fan manufacturers do not include any values for
mechanical components, these noise sources need to be considered in addition to the published fan sound data. Excessive
vibration can also have a noticeable impact on fan noise by
creating resonance with the fan assembly, ducting or mounting
structure. It often acts as an identifier of mechanically generated sound and can be caused by bearing problems, improper
installation, mechanical imbalance of the fan or motor, wear,
fatigue or erosion of the fan or inadequate isolation of the fan
from the mounting structure or duct system.
It is important to understand both the fan sound data published by fan manufacturers and the actual sound realized at the
installed site. Fan manufacturers, in accordance with AMCA
standards, publish fan sound ratings that the user can use to predict expected in situ sound levels. Using eight (8) octave bands
ranging from 45 to 11200 Hz, Fan Sound Pressure (Lp) is measured and recorded at the mid-point frequency of each octave
band. This is conducted in an AMCA certified reverberant or
semi-reverberant room having a calibrated reference sound
source to remove the effects of any environmental and room
noise (called the substitution method). The Fan Sound
Pressures (Lp) are then recorded and converted to Fan Sound
Power (Lw), which then becomes the primary basis for evaluating fan sound. Since Lw is independent of the environment, it
is the only value specific to a particular fan, which makes it
useful in evaluating fan sound levels of different fans. Note that
while Lw is corrected for labratory environmental conditions by
the substitution method, all other environmental and local noise
sources such as motors, drives, bearings, dampers, etc., are not
included in the published fan sound data, as fans are tested
without drives or accessories and with motor sound levels
removed mathematically according to AMCA protocol. And,
since AMCA standards certify Lw values and not Lp values, fan
manufacturers providing sound data in accordance with
AMCA standards will guarantee sound power levels but not
sound pressure levels.
Fan Sound Power (Lw) is normally expressed in decibels
7-26
Industrial Ventilation
(dB) for each octave band and is often converted by the fan
manufacturer to a weighted fan sound pressure value using A
or C scale weighting and recorded as dBA or dBC. However,
since this conversion is only based on A or C scale weighting,
these values do not include any other noise sources. In order to
predict the expected Fan Sound Pressure (dBA or dBC) levels,
the user must convert the Lw levels to the weighted A or C
scale values using both the scale weighting and the effects of
all other noise sources (motor, drive, dampers, environmental
site conditions, etc.). It should be noted that AMCA standards
for sound performance (Lw) allow a +6 dB tolerance for the
first octave band, a +3 dB tolerance for octave bands 2–8 and
an additional +3 dB tolerance, which can be applied to any one
octave band of choice in addition to the previously noted tolerance levels.
Noise Directivity and Distance are important factors in predicting expected fan sound levels. Directivity refers to the
number of radiating surfaces impacting a noise source, while
Distance is simply the straight line distance from the noise
source to the nearest radiating surface. AMCA 303 lists directivities as Q = 1, Q = 2, Q = 4 and Q = 8, with a Q1 directivity
having a spherical form with no reflecting surfaces within the
specified raidus of the sphere, a Q2 directivity having a hemispherical form where one (1) reflective surface is present within the radius of the sphere (fan installed on floor), a Q4 directivity having a quarter sphere form where two (2) reflective
surfaces are present within the radius of the sphere (fan
installed on floor next to wall), and a Q8 directivity having a
one-eighth sphere form where three (3) reflective surfaces are
present within the raidius of the sphere (fan installed on floor
next to adjacent walls). For each additional reflecting surface,
the directivity factor is doubled. When specifying a required
sound presssure level (dBA, dBC), both the distance and the
direction of the control position from the noise source should
be listed, such as 85 dBA @ Directivity Q2 at 5 feet and 60
degrees.
DESIGN OPTIONS FOR NOISE CONTROL
The following noise control options are available and
should be considered by the system designer:
1) Select the fan to operate near the peak total efficiency
and on a stable part of the fan curve.
2) Design the ductwork at the inlet and outlet of the fan
for the best aerodynamics and minimize fan system
effects.
3) When possible, place the fan in a location where the
noise is less objectionable.
4) Add an inlet or outlet duct silencer to control airborne
noise.
5) Add a properly ventilated acoustical enclosure around
the fan and motor assembly to isolate the noise from
the environment.
6) When necessary, use vibration isolation to limit the
transmission of vibration noise to the building foundation or any local structures.
7) Add acoustical flexible duct connectors at the fan inlet
and outlet to prevent transmission of energy to the
ductwork system and break-out noise to the environment.
8) Use a contact type shaft seal to reduce break-out noise
at the shaft entry into the fan housing.
9) TEFC (totally enclosed fan cooled) motors may be
specified for quiet design, silenced with a motor
silencer installed on the cooling fan end of the motor,
or a TENV (totally enclosed non-ventilated) motor can
sometimes be used in lieu of a TEFC motor.
SAFETY AND ACCESSORIES
Safety Guards are always required. Consider all danger
points including the fan inlet, outlet, shaft, drive, and doors.
Construction should comply with all applicable safety requirements. Accessories can help in the installation and future
maintenance requirements. Fan accessories include drains,
cleanout doors, split housings, and shaft seals.
7.3.2 Fan Selection Using Ratings Tables. For a given fan
size, wheel type, flow rate and static pressure, published fan
capacity tables can be used to determine fan outlet velocities,
fan rpm and power. Fan performance data from capacity tables
are based on standard air at 0.075 lbm/cu ft density. For nonstandard air, the user must correct the actual fan static pressure
to an equivalent fan static pressure at 0.075 lbm/cu ft density
using the density factor. Once corrected, the capacity tables
can then be used to determine the fan rpm and power. Since the
published power value is at standard conditions, the density
factor must be applied again to the published power value to
determine the actual operating power. Both power values are
important, as the fan may be required to start at standard conditions and transition to operating conditions as the system
comes on-line.
Capacity tables published by fan manufacturers are based
on either Fan Total or Fan Static Pressure (FTP, FSP). As
defined by AMCA, Fan Total Pressure is the increase in total
pressure from the fan outlet to the fan inlet and Fan Static
Pressure is the Fan Total Pressure minus the velocity pressure
at the fan outlet (VPout).
FTP = TPoutlet – TPinlet
[7.1]
FTP = (SPoutlet + VPoutlet) – (SPinlet + VPinlet)
and,
FSP = FTP – VPout,
FSP = (SPoutlet + VPoutlet) – (SPinlet + VPinlet) – VPout
FSP = SPoutlet – SPinlet – VPinlet
[7.2]
From the above, fan total pressure can also be expressed as
FTP = FSP + VPout
[7.3]
Fans
7.3.3 Fan Efficiency. The most common type of fan capacity table is a “multi-rating table” (Table 7-1), which displays a
range of fan capacities. For a given pressure, the highest efficiency will usually be in the middle third of the column for
flow rate. Some manufacturers show the rating of maximum
efficiency for each pressure with highlights. In the absence of
such a guide, the designer can calculate Fan Total or Static
Efficiency (FTE, FSE) for single stage fans using the following equations:
7-27
is mainly because, as the selection programs become more
capable, some fan manufacturers will publish abbreviated
capacity tables for use as a guide, which creates a greater variance in the interpolation process. And, while published capacity tables are only updated when republished, databases for
electronic selection programs are often continuously updated
for real time accuracy.
Fan Total Efficiency,
ηT = (Q H FTP) ÷ (CF H PWR)
[7.4]
Fan Static Efficiency,
ηs = (Q H FSP) ÷ (CF H PWR)
[7.5]
where:
ηT = Fan Total Efficiency (FTE)
ηs = Fan Static Efficiency (FSE)
Q = Volumetric flow rate, acfm [acms]
FTP = Fan Total Pressure, "wg [kPa, Pa]
FSP = Fan Static Pressure, "wg [kPa, Pa]
PWR = Fan Shaft Horsepower, hp [watts]
CF = Conversion factor, 6343 [1]
Note: Fan Pressure and Power must be the same basis, i.e.,
both are either actual or equivalent values.
Even with a multi-rating table, it is often necessary to interpolate in order to select the fan speed (RPM) and power
(PWR) for the exact conditions desired. In some cases, a double interpolation may be necessary. Straight line interpolations
using the multi-rating table can be used for preliminary design,
but the actual fan performance for the final design conditions
should be confirmed with the fan manufacturer.
Centrifugal fans may be offered in AMCA designated performance Classes I through V (Figure 7-16). A fan designated
for a particular class must be physically capable of operating
at any point within the performance limits for its class.
Performance limits for each class are in terms of outlet velocity and static pressure. Multi-rating tables will also usually be
shaded to indicate the selection zones for various classes and
usually state the maximum operating RPM (at standard conditions). Fan class definitions are based on performance and do
not dictate fan materials of construction.
Many fan manufacturers have electronic programs for generating fan performance data and curves. These programs can be
used to select the fan type, size and options and to calculate performance data including fan speed, power, efficiency and sound.
Typical input data and filters include flow rate, fan static or total
pressure, air stream density at the fan inlet, site elevation, operating and maximum temperatures and sound parameters.
For fan manufacturers having both published fan capacity
tables and electronic selection programs, the most accurate
selection process will normally be the electronic program. This
EXAMPLE PROBLEM 7-1 (Determining Fan Efficiency
(IP Units)
An existing fan is suspected of operating at a low efficiency
and is being evaluated for possible replacement with a higher
efficiency fan for energy savings. The existing fan is a high efficiency radial that was originally installed downstream a cyclone
that has recently been replaced with a baghouse. Field measurements are taken and it is found that the fan capacity is
20000 acfm at 12 "wg FSP, 56 hp and an outlet velocity of 4330
fpm. Determine the fan total and static efficiency.
Using Equation 7.4, fan total efficiency can be calculated as:
Fan Total Efficiency, ηT = (Q H FTP) ÷ (CF H PWR)
Solving for VPout = (4330/4005)2 = 1.17 "wg, then
FTP = FSP + VPout = 12 + 1.17 = 13.17 "wg, and
ηT = (Q H FTP) ÷ (CF H PWR)
ηT = (20000 H 13.17) ÷ (6343 H 56)
ηT = 74.2%
Using Equation 7.5, fan static efficiency can be calculated as:
Fan Static Efficiency, ηs = (Q H FSP) ÷ (CF H PWR)
ηs = (20000 H 12) ÷ (6343 H 56)
ηs = 67.5%
Fan Static Efficiency is often used to determine the efficiency of housed fans and is always used for unhoused fans not
having an outlet to capture the velocity pressure component.
Fan Total Efficiency is used to determine the efficiency of
housed fans, as the velocity pressure component is captured at
the fan outlet. While static pressure is used in system design
calculations to determine the system static pressure, and while
fans are normally selected on the basis of fan static pressure,
fan efficiency can be evaluated using either fan static pressure
(fan static efficiency) or fan total pressure (fan total efficiency), depending on the user’s interest.
7.3.4 Fan Selection at Non-Standard Density. Fan performance is affected by changes in air stream density. If the
cumulative correction to the air stream density for duct pressure, temperature, moisture, elevation and gas composition is
7-28
Industrial Ventilation
TABLE 7-1 (IP). Example of Multi-Rating Table
TABLE 7-1 (SI). Example of Multi-Rating Table
Fans
FIGURE 7-16. AMCA fan classes, airfoil and backward inclined single width (Reprinted from AMCA Publication 99-16,
STANDARDS HANDBOOK by permission from AMCA International)(7.2)
7-29
7-30
Industrial Ventilation
less than 5%, while corrections to flow rate and pressure are
slight they are recommended. However, if the cumulative correction to the air stream density is equal to or greater than 5%,
then corrections must be made. Note that published fan rating
tables and curves are based on standard air density at the fan
inlet, while electronic selection software can often be run for
either standard or actual conditions.
Fan Pressure (P) and Power (PWR) vary directly with
changes in air stream density. Flow rate (ACFM), however, is
effected only by the density factor of the duct pressure at the
fan inlet. The fan selection process requires fan rating tables to
be entered with actual volumetric flow rate and a corrected or
equivalent pressure, although most electronic software programs accommodate either actual or standard operating conditions. In all cases, the designer should examine the fan selection for both standard and actual operating conditions to allow
for variables such as maximum or minimum temperatures and
motor starting and operating requirements for all conditions.
Fan equivalent and actual pressures can be calculated as:
Peqv = Pact ÷ df or
[7.6a]
Peqv = (Pact) (ρstd ÷ ρact) and
Pact = (Peqv) (df) or
[7.6b]
Pact = (Peqv) (ρact ÷ ρstd)
where:
Peqv = Equivalent pressure at standard air density,
"wg [kPa, Pa]
Pact = Actual pressure at actual air density,
"wg [kPa, Pa]
ρstd = Standard air density, 0.075 lbm/ft3
[1.204 kg/m3]
ρact = Actual air density, lbm/ft3 [kg/m3]
df = density factor
Likewise, fan power is calculated for standard and actual conditions as:
FAN-PWRstd = PWRact ÷ df or
[7.7a]
FAN-PWRstd = (PWRact) (ρstd ÷ ρact) and
FAN-PWRact = (PWRstd) (df) or
[7.7b]
FAN-PWRact = (PWRstd) (ρact ÷ ρstd)
where:
FAN-PWR = Fan shaft power, hp [kW, W]
FAN-PWRstd = Fan shaft power at standard air density,
hp [kW, W]
FAN-PWRact = Fan shaft power at actual air density,
hp [kW, W]
ρstd = Standard air density, 0.075 lbm/ft3
[1.204 kg/m3]
ρact = Actual air density, lbm/ft3 [kg/m3]
df = density factor
Pressures (Peqv and Pact) can be fan static or total pressure,
depending on the fan manufacturer’s rating method. A fan
selected from published tables operates at the speed and actual
volumetric flow rate shown in the tables, but the fan pressure
and power will only be the values shown in the tables when the
density factor is 1.0. When the density factor is not 1.0, the fan
pressure and power will be the values shown in the table converted to actual conditions.
Fan selection at non-standard conditions requires knowing the
actual volumetric flow rate at the fan inlet, the actual pressure
requirement (FSPact, FTPact), and the air stream density at the fan
inlet. Determining these variables requires that the system designer account for the effect of density as discussed in Chapter 9.
In some cases, fans used for high temperature applications
must be “cold started.” When the air is cold the air density is
greater and the fan’s motor may be in danger of operating
beyond its horsepower rating. Variable frequency drive (VFD)
speed controls or mechanical controls such as fan dampers
may be useful in these applications. Also, ambient temperatures in very cold climates sometimes provide enough density
change to affect the motor horsepower rating. For example,
fans selected at standard temperature (70 F) [21 C] during winter conditions can have a 20% increase in air density at -20 F
[-6.7 C] ambient. This can increase horsepower requirements
20% over standard conditions.
EXAMPLE PROBLEM 7-2 (Effect of Air Stream Density
on Fan Selection) (IP Units)
The following example compares a fan selection for a 10000
acfm system operating with variations in environmental conditions in Chicago and Houston. The system is designed and the
fan is selected based on standard conditions, but the designer
fails to correct the fan selection to the actual conditions at each
location. The results are:
System design and fan selection based on standard conditions: 10000 acfm at 8 "wg equivalent fan static pressure, 2161
rpm, 16.3 hp, 0.075 lbm/ft3 fan inlet density (df = 1.0).
Chicago Winter Conditions: 20 F, 40% relative humidity,
0.0836 lbm/ft3 fan inlet air stream density (df = 1.12).
Houston Summer Conditions: 80 F, 80% relative humidity,
0.0725 lbm/ft3 fan inlet air stream density (df = 0.97).
Actual Chicago System Performance: 10000 acfm @ 8.96
"wg FSP, 2161 rpm, 18.3 hp at 0.0836 lbm/ft3 fan inlet air
stream density. Note also that the system mass flow rate (volume x density) is 836 lb/min.
Actual Houston System Performance: 10000 acfm @ 7.76
"wg FSP, 2161 rpm, 15.8 hp at 0.0725 lbm/ft3 fan inlet air
stream density. Note also that the system mass flow rate (volume H density) is 725 lb/min.
Fans
Summary: Since fans are constant volume, at a fixed speed
a fan selected for 10000 acfm will move 10000 acfm at all three
(3) conditions. However, since pressure and power vary directly with the change in air density, failing to correct the fan selection based on the actual operating conditions results in a 15%
difference in system operating pressure, power and mass flow
rate between the “identical” installed systems. Using the density factor to correct the fan selection for actual conditions will
ensure that the system as designed will operate as intended at
its installed location.
7-31
at Position 2 as an induced draft fan. Determine the fan selection criteria for each position. Note that when the fan is installed
at Position 1, the entry hood (H) is replaced by an inlet filter
connecting to the fan inlet with a 21" [534 mm] diameter spool
section and having the same pressure loss as the hood. Since
this is a curing system, the effects of moisture are negligible.
Fan Installed at Position 1, Forced Draft Flow:
Flow rate: 4000 acfm at 70 F with df = 1.0 [1.89 acms at 21.1
C with df = 1.0]
Fan Static Pressure: FSP = SPout – SPin – VPin
The static pressure on the discharge of the fan (SPout) is the
sum of the system losses from the fan discharge to Point D and
is 0.47 "wg + 1.29 "wg + 0.66 "wg = 2.42 "wg [117.1 Pa + 321.3
Pa + 164.4 Pa = 602.8 Pa].
The static pressure on the inlet of the fan (SPin) is from the
inlet filter and is -0.83 "wg [206.7 Pa].
EXAMPLE PROBLEM 7-3 (Fan Selection, Curing
Process) (IP Units With SI Unit Conversions)
Figure 7-17 depicts a curing system in which 70 F [21.1 C]
ambient air enters the system, is heated to 600 F [316 C] for
the curing process and is then discharged to the atmosphere.
The air enters the system at Point A and is conveyed through
50 ft [15.2 m] of 15" [381 mm] diameter ducting to Point B. At
Point B the air is heated to 600 F [316 C] and conveyed through
50 ft [15.2 m] of 21" [534 mm] diameter ducting with two 90
degree elbows to the curing chamber at Point C. From Point C
the air is conveyed through 50 ft [15.2 m] of 21" [534 mm] diameter ducting to Point D, where it is released to the atmosphere.
The fan can be installed at Position 1 as a forced draft fan or
FIGURE 7-17. Curing process
The velocity pressure on the inlet side of the fan (VPin) is the
velocity pressure of the air in the 21" [534 mm] diameter spool
connecting the inlet filter to the fan and is 0.17 "wg [42.3 Pa].
Then,
FSP = SPout – SPin – VPin
FSP = 2.42 – (-0.83) – 0.17 = 3.08 "wg
[602.8 – (-206.7) – 42.3 = 767.2 Pa]
The fan selection criteria for Fan Position 1 is 4000 acfm @
3.08 "wg FSP, 70 F with df = 1.0 [1.89 acms @ 767.2 Pa FSP,
21.1 C with df = 1.0].
Fan Installed at Position 2, Induced Draft Flow:
Flow rate: The flow rate entering the system at Point A is
4000 acfm at 70 F with df1 = 1.0 [1.89 acms at 21.1 C with df1
7-32
Industrial Ventilation
= 1.0]. However, when the air passes through the heater at
Point B, the air volume changes directly with the ratio of the air
stream density factors, df1/df2. Since the density factor for 600
F [316 C] is 0.5 (Chapter 9), the flow rate required at the fan
inlet is:
Q2 = (Q1) (df1/df2) = (4000) (1.0/0.5) = 8000 acfm
(1.89 acms H 767.2 Pa) H (1 + 0.09) = 2107 watts
(1.0) (0.75)
Fan Position 2: MTR-PWRoutput-act =
(Q H FSP) H (1 + Fdrv)
(6343) (FSE)
[IP]
MTR-PWRoutput-act =
[(1.89) (1.0/0.5) = 3.78 acms]
The flow rate at Fan Position 2 is then 8000 acfm @ 600 F
with df2 = 0.5 [3.78 acms @ 316 C with df2 = 0.5].
(8000 acfm H 2.90 in wg) H (1 + 0.06) = 5.17 hp
(6343) (0.75)
(Q H FSP) H (1 + Fdrv)
(1.0) (FSE)
Then, calculating the Fan Static Pressure:
FSP = SPout – SPin – VPin
Since the fan discharges to atmosphere using a no-loss
stack, the static pressure on the fan discharge (SPout) is 0.0 "wg.
The static pressure on the inlet of the fan (SPin) is the sum
of the system losses from Point A to Point D and is (-1.30 "wg)
+ (-1.29 "wg) + (-0.66 "wg) = -3.25 "wg [-323.8 Pa + (-321.3 Pa)
+ (-164.4 Pa = 809.5 Pa].
The velocity pressure on the inlet side of the fan (VPin) is the
velocity pressure of the air in the 21" [534 mm] diameter duct
at the fan inlet and is
VP = (V/4005)2 (df2) = (3326/4005)2 (0.5) = 0.35 "wg
[SI]
MTR-PWRoutput-act =
(3.78 acms H 722.7 Pa) H (1 + 0.08) = 3934 watts
(1.0) (0.75)
Since the fan selected at Position 2 is operating at 600 F
[316 C], its power is corrected to standard conditions using
Equation 7.16 as follows:
MTR-PWRoutput-std = (MTR-PWRoutput-act) ÷ df
MTR-PWRoutput-std = 5.17 hp ÷ 0.5 = 10.3 hp
[3934 ÷ 0.5 = 7868 watts]
The forced draft fan at Position 1 will operate at 2.77 hp (df
= 1.0), while the induced draft fan at Position 2 will operate at
5.17 hp (df = 0.5). If the induced draft fan at Position 2 is operated with an air stream lower than 600 F, the motor should be
sized accordingly.
[(V/1.29)2 (df2) = (17.0/1.29)2 (0.5) = 86.8 Pa]
then, FSP = SPout – SPin – VPin
FSP = 0.0 – (-3.25) – 0.35 = 2.90 "wg
[0.0 – (-809.5) – 86.8 = 722.7 Pa]
The fan selection criteria for Position 2 is 8000 acfm @ 2.90
"wg FSP, 600 F with df2 = 0.5 and 8000 acfm @ 5.80 "wg
equivalent FSP at 70 F with df1 = 1.0 [3.78 acms @ 722.7 Pa
FSP, 316 C with df2 = 0.5 and 3.78 acms @ 1445.4 Pa equivalent FSP at 21.1 C with df1 = 1.0].
Comparing Fan Selections for Positions 1 and 2:
Fan Size: Based on the volumetric flow rates, an 18" [457
mm] fan would likely be selected for Position 1 and a 24" [610
mm] fan for Position 2. This means that placing the fan at
Position 1 would result in lower capital cost and a smaller
space requirement.
Motor Power: Using Equation 7.15 from Section 7.4.7,
Figure 7-15 and an estimated fan static efficiency of 75%, the
motor horsepower for each fan can be approximated as follows:
Fan Volumetric and Mass Flow Rate Comparison:
Mass flow rate (lb/min, kg/m3) is the volumetric flow rate
times the air density corrected by the density factor.
Fan Position 1 is (4000 acfm) (0.075 lb/ft3) (1.0) = 300
lbm/min [(1.89 acms) (1.204 kg/m3) (1.0) = 2.28 kg/s]. Fan
Position 2 is (8000 acfm) (0.075 lb/min) (0.5) = 300 lbm/min
[(3.78 acms) (1.204 kg/m3) (0.5) = 2.28 kg/s]. While the fan volumetric flow rates are different for the two conditions, their
mass flow rates are the same, as systems are normally
designed for constant mass flow and fans are selected for the
actual volumetric flow rate at the fan inlet.
From these comparisons, the most economical fan for this
application based on capital cost, size and power is the smaller
fan at Position 1.
Fan Position 1: MTR-PWRoutput-act =
(Q H FSP) H (1 + Fdrv)
(6343) (FSE)
[IP]
MTR-PWRoutput-act =
(4000 acfm H 3.08 in wg ) H (1 + 0.07) = 2.77 hp
(6343) (0.75)
(Q H FSP) H (1 + Fdrv)
(1.0) (FSE)
MTR-PWRoutput-act =
7.4
[SI]
FAN AND SYSTEM PERFORMANCE
A well designed ventilation system is one in which the
specified Point of Operation on the system curve intersects the
fan curve at a stable point of operation. As such, it is important
to understand the characteristics of both fan and system curves.
Fans
7.4.1. Point of Operation. Fans are usually selected for
operation on a fan curve at a specified condition called the
point of operation. Both fans and systems have variable performance characteristics that can be graphically represented as
curves depicting an array of operating points. The actual point
of operation will be the single point at the intersection of the
system curve and the fan curve.
FAN PERFORMANCE CURVE
Fan performance curves are plotted to represent a fan’s
operating characteristics. Figure 7-18 is a typical graph where
flow rate (Q) is on the x-axis and pressure (P) and power
(PWR) are plotted on the y-axis. Fan speed (RPM), fan wheel
diameter (d) and air stream density (ρ) at the fan inlet are constant and should always be clearly stated. Other variables such
as total or static efficiency may also be included on the curve.
Figure 7-18 shows that the maximum airflow occurs when
there is no resistance to flow and that no airflow occurs when
the fan inlet or outlet is blocked. Pt is plotted as total pressure
and Ps as static pressure. Since total pressure = static pressure
+ velocity pressure, the range between the individual points of
operation on curves Pt and Ps is the velocity pressure. Note that
the exact shape of the fan curve depends on the fan design.
SYSTEM CURVE
Every system has a resistance to flow from the individual
components in the system. Figure 7-19 illustrates the variations of pressure (P) with flow rate (Q) for three typical situations.
In turbulent flow, Pressure (P) varies as the square of the
flow rate (Q). This is commonly found in mechanical components of a system such as ducting and fittings. In laminar flow,
Pressure (P) varies directly with the change in flow rate (Q).
This is often found in low pressure dynamic system components such as low velocity air filters. In constant pressure,
Pressure (P) is constant over a range of flow rates (Q). This is
often found in some types of wet scrubbers and fluidized beds.
In some dynamic system components such as dust collectors
or packed towers, the media may have a pressure relationship
to flow that does not follow any of these examples and is
instead a numerical value from a flow-pressure chart, a laboratory test result, or a field measurement. In these cases, the
equipment or media supplier should be consulted (see
Example Problems 7-5A and 7-5B).
For a fixed system, the system curve is the result of the combined effects from the individual components. When all system components are in turbulent flow, the system is called a
Turbulent Flow System and the points of operation of the system curve are calculated with the system pressure varying with
the square of change in flow rate. When one or more of the
system components is operating in non-turbulent flow conditions, the system is called a Hybrid Flow System and each
point of operation of the system curve is calculated as the sum
7-33
of the turbulent and non-turbulent pressures at each corresponding flow rate. In hybrid flow systems, when the effects
of the non-turbulent flow conditions are insignificant relative
to the turbulent flow conditions they can be ignored. In these
cases the hybrid flow system can be treated as a turbulent flow
system.
Typical plots of system pressure and flow rate for three different and arbitrary fixed duct systems (Systems A, B and C)
are illustrated in Figure 7-20. For a fixed system, a change in
flow rate results in a change in system pressure along the system curve. If the system components change, the system resistance changes and the shape of the system curve also changes.
For example, with a system operating in turbulent flow conditions at the design flow rate (Q) and at the design system pressure (P), an increase in flow rate to 120% of Q will result in an
increase in system pressure (P) of 144%. Likewise, a decrease
in flow rate Q to 50% would result in a decrease in system
pressure P to 25% of the design system pressure. In Figure 720, System Curve B is representative of a system that has a
higher loss than System Curve A and System Curve C has a
lower loss than System Curve A. The system design point is
the point on the system curve at which the fan is to be selected
to provide the flow rate and pressure requirements of the system.
7.4.2 Matching the Fan Curve and the System Curve.
The design point of operation results from the process of
designing a system and selecting a fan. The point of intersection of the system curve and the fan curve determines the point
of operation. Figure 7-21 depicts the intersection of the fan
curve and the system curve at the point of operation (A) in
addition to conditions that can result from a poor system
design and fan selection process (B, C and D).
Figure 7-22 depicts the range in performance of a welldesigned system and fan selection based on a ± 10% system
design tolerance and the AMCA certified fan rating tolerance.
Fan manufacturers offering either AMCA Certified Fans for
Air Performance or fans that are certified by the fan manufacturer according to AMCA Standards for Air Performance will
generally have a tolerance for flow rate and pressure of -3%
between 20–60% of free delivery (and higher outside of this
range), while the tolerance for power is +5% from free delivery to shut off. This means that the selected fan could deviate
from the design point of operation by these margins and still
be within its range of certification. Note also that AMCA
Certification for Air Performance is normally for a specified
range of the fan curve and not from full open to shut off. As
such, the fan should be selected for the point of operation to lie
on the certified portion of the fan curve.
From Chapter 9, a designer determines the calculated volumetric flow rate (Q) and system static pressure (SSP), shown
as Point A in Figure 7-23. However, the designer typically
adds a safety factor for volume and/or pressure. This is also
shown in Figure 7-23 as Point B and will be found up and to
the right of Point A. What we would expect to find in the field
7-34
Industrial Ventilation
FIGURE 7-18. Typical fan performance curve
Fans
7-35
operation at some point other than the design point of operation. When this occurs, it may become necessary to alter the
system. Because the fan performance curve is specific to a
given fan at a designated speed (RPM), a change of fan speed
can be relatively simple if a belt drive arrangement is used or
if the fan is operating on a VFD controller. The “Fan Affinity
Laws” (Section 7.4.4) are useful when changes in fan performance are required.
7.4.3 Affinity Laws for Fans and Systems. The Affinity
Laws are used to calculate additional points of operation for
both fan and system curves. When applied to fan curves, they
are referred to as the Fan Affinity Laws, and when applied to
system curves, the System Affinity Laws. An understanding of
both the Fan Affinity Laws and the System Affinity Laws is
important to accurately match the point of operation between
the fan curve and the system curve.
7.4.4 Fan Affinity Laws Applied to Fan Curves. The Fan
Affinity Laws (Fan Laws) are used to either determine additional points of operation between two or more fan curves
using the same fan, or to predict a corresponding point of operation between different fan sizes within a homologous fan
series (see Figures 7-24 and 7-26).
DETERMINING ADDITIONAL POINTS OF OPERATION
FOR FAN CURVES USING THE SAME FAN
To determine additional points of operation for fan curves
using the same fan, the Fan Affinity Laws are used as follows:
Flow Rate: Since a fan wheel has a fixed volumetric capacity, volumetric flow rate varies directly with the change in fan
speed and is expressed as:
Q2/Q1 = RPM2/RPM1, or
Q2 = (Q1) (RPM2/RPM1)
[7.8a]
If the flow rate is known, this can be rewritten to solve for
speed as:
RPM2/RPM1 = Q2/Q1, or
RPM2 = (RPM1) (Q2/Q1)
Pressure: Since fans are always in turbulent flow, pressure
varies directly with the change in air density and the square of
the change in either flow rate (Q) or fan speed (RPM) and is
expressed as:
FIGURE 7-19. System curves
P2/P1 = (Q2/Q1)2 (df2/df1), or
P2 = (P1) (Q2/Q1)2 (df2/df1)
[7.9a]
If fan speed (RPM) is known, this can be rewritten as:
would be the intersection of the pressure curve selected for
Point B and the real system curve, a third point, Point C. Point
C will be the system point of operation except for changes to
system components (hoods, ducting), damper settings, types
of filter media, plugged filters, etc., which would change the
system curve.
There are a number of reasons why the system design, fan
selection, fabrication, and installation process can result in
P2/P1 = (RPM2/RPM1)2 (df2/df1), or
P2 = (P1) (RPM2/RPM1)2 (df2/df1)
Power: Since fans are always in turbulent flow, power
varies directly with the change in air stream density and the
cube of the change in either flow rate (Q) or fan speed (RPM)
and is expressed as:
PWR2/PWR1 = (Q2/Q1)3 (df2/df1), or
PWR2 = (PWR1) (Q2/Q1)3 (df2/df1)
[7.10a]
7-36
Industrial Ventilation
FIGURE 7-20. System curves
If fan speed is known, this can be rewritten as:
PWR2/PWR1 = (RPM2/RPM1)3 (df2/df1), or
PWR2 = (PWR1) (RPM2/RPM1)3 (df2/df1)
(Speed increases 10%)
P2 = (P1) (Q2/Q1)2 (df2/df1)
= (10) (13200/12000)2 (1.0) = 12.1 "wg
[2.49 H (6.23/5.66)2 (1.0) = 3.0 kPa]
(Pressure increases 21%)
PWR2 = (PWR1) (Q2/Q1)3 (df2/df1)
EXAMPLE PROBLEM 7-4 (Applying the Fan Affinity Laws
to the Same Fan) (IP Units With SI Unit Conversions)
A fan is operating at 12000 acfm (Q1), 10 "wg FSP (P1) and
25 hp (PWR1) [5.66 acms, 2.49 kPa, 18.6 kW]. The fan speed
is 1000 rpm (RPM1) with df = 1.0. Find the new point of operation on a fan curve for a 10% increase in fan capacity.
Q2 = (Q1) (1.1)
= (12000 acfm) H (1.1) = 13200 acfm
[(5.66 acms) H (1.1) = 6.23 acms]
(Flow rate increases 10%)
RPM2 = (RPM1) (Q2/Q1)
= (1000) (13200/12000) = 1100 rpm
[(1000) (6.23/5.66) = 1100 rpm]
= (25)(13200/12000)3(1.0) = 33.3 hp
[(18.6) (6.23/5.66)3 (1.0) = 24.8 kW]
(Power increases 33%)
The point of operation for a new fan curve for Condition 2 is
13200 acfm @ 12.1 "wg FSP, 33.3 hp, 1100 rpm with df = 1.0
[6.23 acms @ 3.0 kPa, 24.8 kW, 1100 rpm with df = 1.0].
Additional points of operation can be calculated from Condition
1 for Q3, Q4, Q5, etc. as required.
Applying the Fan Affinity Laws to a specific fan allows accurate projections of additional points of operation on a new fan
curve based on a known flow rate, speed, pressure, power and
air stream density.
Fans
7-37
7-38
Industrial Ventilation
FIGURE 7-22. AMCA Air Performance with Certified Rating Tolerance (Reprinted from AMCA Publication 200-95 (R2011), Air
Systems by permission of AMCA International(7.5)
DETERMINING ADDITIONAL POINTS OF OPERATION
FOR A FAN CURVE BETWEEN DIFFERENT FAN SIZES
The Fan Affinity Laws can be expanded to predict the point
of operation on a fan curve of a different sized fan as long as
both fans are within a homologous fan series. A homologous
fan series represents a range of fan sizes in which all fan air
stream dimensions are geometrically proportional (Figure 726).
Homologous fans that are AMCA compliant must be
designed to achieve complete geometric, kinematic and
dynamic similarities according to published AMCA standards.
Additionally, AMCA standards only allow scaling to a larger,
and not a smaller fan size.
If the designer applies the Fan Affinity Laws to a homologous fan series not complying with AMCA standards for similarity, the deviations may be unpredictable, and in these cases
the fan manufacturer should be consulted to determine acceptable results. If the designer applies the Fan Affinity Laws to a
homologous fan series complying with AMCA standards for
similarity and scales up and not down, only minor deviations
having an insignificant impact would be expected.
When applying the Fan Affinity Laws across fan sizes, variables for wheel diameter (d) and fan efficiency (η) are introduced. Since changes in efficiency are normally insignificant,
the change in wheel diameter becomes the most important
variable. The Fan Affinity Laws shown below are expanded to
be used across homologous fan sizes.
Flow Rate: Since a fan wheel has a fixed volumetric capacity, its volumetric flow rate varies directly with the change in
fan speed and exponentially with the change in wheel diameter.
If speed is known, the flow rate can be calculated by:
Q2/Q1 = (RPM2/RPM1) (d2/d1)3 or
Q2 = (Q1) (RPM2/RPM1) (d2/d1)3
[7.8b]
If the flow rate is known, speed can be calculated by:
RPM2/RPM1 = (Q2/Q1) (d2/d1)-3 or
RPM2 = (RPM1) (Q2/Q1) (d2/d1)-3
Pressure: Since fans are always in turbulent flow, pressure
varies directly with the change in air density and with the
square of the change in either flow rate (Q) or fan speed
(RPM) and exponentially with the change in wheel diameter,
and is expressed as follows:
If the flow rate is known, pressure can be calculated by:
P2/P1 = (Q2/Q1)2 (d2/d1)-4 (df2/df1), or
P2 = (P1) (Q2/Q1)2 (d2/d1)-4 (df2/df1)
If speed is known, pressure can be calculated by:
[7.9b]
Fans
7-39
FIGURE 7-23. Fan selection at standard conditions
P2/P1 = (RPM2/RPM1)2 (d2/d1)2 (df2/df1), or
P2 = (P1) (RPM2/RPM1)2 (d2/d1)2 (df2/df1)
Power: Since fans are always in turbulent flow, power
varies directly with the change in air density, with the cube of
the change in either flow rate (Q) or fan speed (RPM), and
exponentially with the change in wheel diameter and is
expressed as follows:
Note: The Fan Affinity Laws are for incompressible flow
and assume that fan efficiency is constant or changes are
insignificant. Values for air density (ρ) can be substituted for
the density factor (df) if the designer so chooses.
PWR2/PWR1 = (RPM2/RPM1)3 (d2/d1)5 (df2/df1) or
In practice, the Fan Affinity Laws are most often applied to
a single fan size to determine the effect of changing only one
variable. A fan performance curve is always specific to a fan
of a given size operating at a single speed (RPM). If the fan
speed is increased, it is represented by a new fan curve that
will move up and to the right of the original fan curve depicted
as points “1” and “2” in Figure 7-24. When the fan speed is
decreased, the opposite effect occurs. The relationship
between Q and P is a family of fan curves for different fan
speeds.
PWR2= (PWR1) (RPM2/RPM1)3 (d2/d1)5 (df2/df1)
Changes in Air Density: System static pressure, fan static
If the flow rate is known, power can be calculated by:
PWR2/PWR1 = (Q2/Q1)3 (d2/d1)-4 (df2/df1) or
PWR2 = (PWR1) (Q2/Q1)3 (d2/d1)-4 (df2/df1) [7.10b]
If the speed is known, this can be rewritten as:
7-40
Industrial Ventilation
FIGURE 7-24. Effect of 10% increase in fan speed
pressure and fan power are based on the density of the air at
either the inlet duct to the fan (SSP) or at the fan inlet (FSP).
Figure 7-25 illustrates the effect of density variation on the
system curve, fan curve and fan power. Both pressure and
power vary directly with the change in air density or the density factor.
7.4.5 Limitations on the Use of the Fan Affinity Laws.
Fan Affinity Laws are equations used to predict alternate
points of operation for fans operating in turbulent flow conditions and at the same relative points of operation on each fan
curve. Care must be taken to apply the fan affinity laws
between the same relative points of operation. Figure 7-26 is a
typical representation of two homologous fan curves, FAN-1
and FAN-2 operating at RPM1 and RPM2, respectively.
Assuming a point of rating indicated as A1 on FAN-1, there is
only one location on FAN-2 with the same relative point of rating and that is at A2. The Fan Affinity Laws can be used to
identify every other point that would have the same relative
point of rating as A1 and A2 along the same system curve. The
curves shown representing the same relative points of rating in
Figure 7-26 are system curves for turbulent flow conditions.
Care must be exercised when applying the Fan Affinity
Laws in the following cases:
1. Where any system component does not operate in turbulent flow (e.g., filter media losses).
2. Where the system has been physically altered or for
any other reason operates on a different system curve.
7.4.6 The System Affinity Laws Applied to System
Curves. While the Fan Affinity Laws are used to determine
additional points of operation for flow rate and pressure on the
fan curve, the System Affinity Laws are used to determine
additional points of operation for flow rate and pressure on the
system curve. This is shown in Figure 7-26, where a new point
of operation is plotted on system curve “A” from A1 to A2, and
on system curve “B” from B1 to B2.
To determine additional points of operation for a system
curve using a volumetric flow rate and system static pressure
from the calculation worksheet, the system designer would use
the following guidelines depending on whether or not the system is operating in turbulent or hybrid flow conditions.
Turbulent Flow Systems are systems operating wholly in turbulent flow, while Hybrid Flow Systems are systems having
one or more system component operating in non-turbulent
flow conditions.
Fans
FIGURE 7-25. Effect of 50% decrease in gas density
7-41
7-42
Industrial Ventilation
square root of change in density and is expressed as:
Q2/Q1 = [(SSP2/SSP1)0.5 (df2/df1)-0.5 ] or
Q2 = (Q1) [(SSP2/SSP1)0.5 (df2/df1)-0.5]
[7.12]
Calculating the values for Q2, Q3, Q4, etc., and SSP2, SSP3,
SSP4, etc., the designer can construct the system curve for a
system operating in turbulent flow conditions.
EXAMPLE PROBLEM 7-5A (Using the System Affinity
Laws to Calculate Additional Points of Operation on a
System Curve in Turbulent Flow; Increase Flow 10%) (IP
Units with SI Unit Conversions)
A system is designed to operate at 15000 acfm and 10 "wg
system static pressure with df = 1.0 [7.08 acms, 2.49 kPa]. The
system is operating in fully turbulent flow conditions and the
designer wants to allow for a 10% future increase in system
flow rate. Find the capacity for the future expansion.
FIGURE 7-26. Homologous fan performance curves
DETERMINING ADDITIONAL POINTS OF OPERATION
FOR A SYSTEM CURVE IN TURBULENT FLOW
Since fans always operate in turbulent flow, then by substituting system static pressure (SSP) for pressure (P), the
System Affinity Laws used for turbulent flow systems are
identical to the Fan Affinity Laws for flow and pressure.
Knowing the system volumetric flow rate (Q), the system
static pressure (SSP), and the density factor (df) from the calculation worksheet, the designer can use the System Affinity
Laws to develop a system curve specific to the system design
as follows:
Using the Flow Basis to Determine Pressure: To determine
new system static pressures corresponding to target flow rates,
first select the target system flow rates, Q2, Q3, Q4, etc., and
then calculate the corresponding system static pressures to plot
the system curve.
Pressure (SSP): System Static Pressure varies directly with
the change in air density and the square of the change in flow
rate (Q):
SSP2/SSP1 = (Q2/Q1)2 (df2/df1) or
SSP2 = (SSP1) (Q2/Q1)2 (df2/df1)
[7.11]
Using the Pressure Basis to Determine Flow: To determine
new system flow rates corresponding to target system static
pressures, first select the target system static pressures, SSP2,
SSP3, SSP4, etc., and then calculate the corresponding flow
rates to determine the flow rates and pressures for plotting the
system curve.
Flow Rate (Q): System flow rate varies with the square root
of the change in pressure (SSP) and the reciprocal of the
Since the system is operating in turbulent flow conditions,
the System Affinity Laws can be used to directly calculate the
future conditions as follows:
Flow Rate:
Q2 = (Q1) (1.1)
= (15000 acfm) H (1.1) = 16500 acfm
[(7.08) (1.1) = 7.8 acms]
(Flow rate increases 10%)
Pressure
SSP2 = (SSP1) (Q2/Q1)2 (df2/df1)
= (10) (16500/15000)2 (1.0) = 12.1 "wg
[(2.49) (7.8/7.08)2 (1.0) = 3.0 kPa]
(Pressure increases 21%)
The system design capacity is 15000 acfm at 10 "wg SSP with
df = 1.0 [7.08 acms at 2.49 kPa with df = 1.0] and a future
capacity to 16500 acfm at 12.1 "wg SSP with df = 1.0 [7.8 acms
at 3.0 kPa with df = 1.0]. The fan and motor should be selected
based on the future capacity with provisions to operate the fan
at the current design.
Fans
DETERMINING ADDITIONAL POINTS OF OPERATION
FOR A SYSTEM CURVE IN HYBRID FLOW
For a Hybrid Flow System, the designer would use the
flow basis noted above and select the target system flow rates
Q2, Q3, Q4, etc. The corresponding system static pressures
(SSP2, SSP3, SSP4, etc.) are a hybrid calculation using the
Affinity Laws for the portion of the system operating in turbulent flow plus a separate calculation for the portion of the system operating in non-turbulent flow. (Note that in many hybrid
flow systems, the effects of non-turbulent flow conditions may
be insignificant and may be ignored at the designer’s discretion).
The new system static pressure for a hybrid flow system is
calculated using Equation 7.13. The new system static pressure for the portion of the system operating in non-turbulent
flow is specific to the particular type of system component(s)
and could vary as a factor of its original pressure, exponentially of its original pressure, or it could be a numerical value or
percentage from a published chart, labratory test, field measurement, etc. For these reasons, it should always be calculated
separately to determine the new system static pressure as follows:
SSPhyb-2 =
[7.13]
[(SSPhyb-1 – |Pnon-trb-1|) (Q2/Q1)2 + |Pnon-trb-2|]
(df2/df1)
where:
SSPhyb-x = the system static pressure of a system
operating in hybrid flow
Pnon-trb-x = the static pressure of the portion of a
hybrid system operating in non-turbulent
flow
7-43
increased flow capacity on the filter media pressure loss and
advises that the new pressure loss for a 10% increased flow
rate through the filter media is 4.4 "wg [1.1 kPa]. The capacity
for the future expansion can now be calculated as:
Flow Rate
Q2 = (Q1) (1.1) = (15000 acfm) (1.1)
[(7.08 acms) (1.1)] = 16500 acfm [7.8 acms]
(Flow rate increases 10%)
Pressure
SSPhyb-2 =
[(SSPhyb-1 – |Pnon-trb-1|) (Q2/Q1)2 + |Pnon-trb-2|]
(df2/df1)
For the portion of the system operating in turbulent flow,
(SSPhyb-1 – |Pnon-trb-1|) (Q2/Q1)2
= (14 – |-4|) (16500/15000)2
[(3,49 – |-1,0|) (7,8/7,08)2]
= 12.1 "wg [3.0 kPa]
(Pressure increases 21% for turbulent flow)
For the portion of the system operating in non-turbulent flow,
|Pnon-trb-2| = |-4.4| = 4.4 "wg [(1.1 kPa]
(Pressure increases 10% in non-turbulent flow)
then
SSPhyb-2 =
[(SSPhyb-1 – |Pnon-trb-1|) (Q2/Q1)2 +
|Pnon-trb-2|] (df2/df1)
= [(14.0 – |-4|) (16500/15000)2 + |-4.4|] (1.0)
[3.49 – |-1.0|) (7.8/7.08)2 + |-1.1|] (1.0)
EXAMPLE PROBLEM 7-5B (Using the System Affinity
Laws to Calculate Additional Points of Operation on a
System Curve in Hybrid Flow; Increase Flow 10%) (IP
Units with SI Unit Conversions)
The system in Example Problem 7-5A is redesigned to
replace a cyclone with a baghouse. The system is now
designed to operate at 15000 acfm at 14 "wg SSP with df = 1.0
[7.08 acms, 3.49 kPa]. Due to non-turbulent flow conditions
through the baghouse filter media, the system is operating in
hybrid flow with 10 "wg [2.49 kPa] representing system turbulent flow conditions and 4 "wg [1.0 kPa] representing non-turbulent flow conditions from the pressure loss through the filter
media. The designer wants to allow for a 10% future increase
in system flow rate. Find the new flow rate and pressure for the
future expansion.
Since the system is operating in hybrid flow, Equation 7.13
is used to calculate the new system static pressure. The filter
media supplier is consulted to determine the effect of the
= 16.5 "wg [4.1 kPa]
(SSP increases 18%)
Summary: Since the system in Example 7-5A is in full turbulent flow, the System Affinity Laws can be used to directly predict additional points of operation along the system curve.
However, since the system in Example 7-5B is in hybrid flow,
the system static pressure is a hybrid calculation using the
Affinity Laws for the portion of the system operating in turbulent
flow and the new pressure loss for the portion of the system
operating in non-turbulent flow. The sum of these values is the
new predicted system static pressure.
The fan and motor should be selected based on future
capacity with provisions to operate the fan at the current
design. However, care must be taken to select the fan and
motor for the future conditions without unnecessarily oversizing
the fan and motor.
7-44
Industrial Ventilation
7.4.7 Correlating System Static Pressure and Fan Static
Pressure to Power. Fans for industrial ventilation systems are
MTR-PWRoutput-act =
specified and selected on the basis of flow rate, static pressure,
and air density. System flow rate, Q, is always stated in actual
cubic feet per minute (ACFM) [actual cubic meters per second, ACMS]. System static pressure, SSP, is stated in inches of
water gauge ("wg) [kPa, Pa]. The air density (ρ) is the density
at the fan inlet. These three (3) components can be used to calculate an approximate system energy requirement and to specify the fan. Fan static pressure (FSP) is stated in inches of
water gauge ("wg) [kPa, Pa] and is used for fan selection and
calculating the final system energy requirement. The density is
always the air density at the fan inlet.
(Flow Rate H Fan Static Pressure) H (1 + Fdrv)
(6343 H Fan Static Efficiency)
System static pressure is used for specifying fan performance and is calculated by the designer prior to fan selection
using the static pressures at the fan inlet and outlet ducts and
the velocity pressure at the fan inlet duct. If the selected fan
has an inlet diameter that is the same size as the system inlet
duct, then the system static pressure will be equal to the fan
static pressure, and they can be used interchangeably. If the
selected fan has an inlet diameter that is a different size than
the system inlet duct, there will be a deviation between the system static pressure and the fan static pressure, and the designer
will need to revise the system design to include a contracting
or expanding fan inlet duct transition fitting, recalculate the
system loss at the inlet of the fan, and then correct the system
static pressure for the fitting loss and the actual velocity pressure at the fan inlet. The revised system static pressure and fan
static pressures are then equal and can be used interchangeably.
In the same way, the energy required by the system can have
a similar deviation if the energy calculation is performed based
on system static pressure prior to fan selection. In most cases,
the designer can use the system static pressure and an estimated fan static or total efficiency (FSE, FTE) prior to fan selection in order to calculate a close approximation of the required
system energy using the following Equations:
MTR-PWRoutput-act =
[7.14] [IP]
(Flow Rate H System Static Pressure) H (1 + Fdrv)
(6343 H Fan Static Efficiency)
MTR-PWRoutput-act =
[7.14] [SI]
(Flow Rate H System Static Pressure) H (1 + Fdrv)
(1.0 H Fan Static Efficiency)
Once the fan is selected and the velocity pressure at the fan
inlet and fan efficiency is known, the designer can accurately
calculate the required system energy using either of the following Equations:
MTR-PWRoutput-act =
[7.15] [IP]
[7.15] [SI]
(Flow Rate H Fan Static Pressure) H (1 + Fdrv)
(1.0 H Fan Static Efficiency)
If the system fan is rated for Fan Total Pressure (FTP), substitute fan total pressure and fan total efficiency in the above
equations.
Once the actual motor operating horsepower is known, then
the motor horespower for standard conditions can be determined by
MTR-PWRoutput-std = (MTR-PWRoutput-act) ÷ df
[7.16]
Energy is referenced as motor horsepower and can be classifed in terms of either motor input power (MTR-PWRinput) or
motor output power (MTR-PWRoutput). Motor input power is
the input power required by the motor from the facility power
supply to energize and drive the motor, and is the sum of the
motor output power, the motor inefficiency, the electrical drive
and controls inefficiency, and any power supply line or transmission losses. Motor output power is the sum of the output
power required at the fan shaft by the system, the fan inefficiency, and the transmission losses between the fan and the
motor. Due to additional losses on the input side of the motor
(electrical drives, controls, etc.), the motor input power will
always be equal to or greater than the motor output power.
For the system designer, it is important to clearly identify
the motor output power to both properly select and size the fan
motor and to ensure that the supply line side power is sized to
provide sufficient input power to the motor. Since the energy
requirement from industrial ventilation systems occurs on the
output side of the motor, the designer is tasked with clearly
identifying motor output power requirements while the task of
identifying the motor input power will be left to those responsible for the facility power supply.
To correctly size the fan motor, the motor output power
should always be calculated on the basis of both actual and standard conditions. Calculations for actual conditions represent the
system operating power, while calculations for standard conditions are used to ensure that the motor output power is sufficient
for starting conditions at standard air stream density.
Fans
7-45
EXAMPLE PROBLEM 7-6A (Design Phase – Approximate
System Energy Calculations Using System Static
Pressure) (IP Units With SI Unit Conversions)
EXAMPLE PROBLEM 7-6B (Final Design – System
Energy Calculations Using the Selected Fan) (IP Units
With SI Unit Conversions)
A system is designed for 10000 acfm of dry air @ 7.43 "wg
system static pressure [4.7 acms, 1.85 kPa], 300 F [149 C] and
2000 ft asl [610 m asl]. The static pressure at the inlet of the fan
is -7.0 "wg [-1.74 kPa] and the static pressure at the outlet of
the fan is +1.0 "wg [+0.25 kPa]. From Chapter 9, the density
factor for temperature and elevation is 0.65, and from Chapter
3, the density factor for -7.0 "wg [-1.74 kPa] duct pressure at
the fan inlet is 0.98, resulting in a density factor at the fan inlet
of 0.65 H 0.98 = 0.64. The inlet ducting to the fan is 22" diameter [559 mm] with a duct velocity of 3788 fpm [19.2 m/s]. The
velocity pressure at the fan inlet ducting is 0.57 "wg [0.14 kPa].
Since the fan has not yet been selected, assume a fan static
efficiency of 72%. The system energy can be approximated as:
A fan is selected for the system in Example 7-6A that has a
20" diameter inlet [508 mm] with inlet and outlet velocities of
4584 fpm [23.3 m/s], a fan static efficiency of 72%, and a fan
total efficiency of 74%. Since the system inlet ducting is 22"
diameter [559 mm], in order to avoid additional losses due to
fan system effects the system designer will need to revise the
system design calculations for a fan inlet duct transition loss
and a revised fan inlet duct velocity pressure, and then calculate the final system energy requirement. Doing so will show
that the revised system static pressure is now equal to the fan
static pressure since the fan inlet duct transition loss is
accounted for. The inlet duct velocity pressure is now equal to
the fan inlet velocity pressure.
MTR-PWRoutput-act =
(Flow Rate H System Static Pressure) H (1 + Fdrv)
(6343 H Fan Static Efficiency)
= (10000 H 7.43) H (1 + 0.045) = 17.0 hp
(6343 H 0.72)
[IP]
= (4.7 H 1.85) H (1 + 0.06) = 12.8 kW
(1.0 H 0.72)
[SI]
and
From Chapter 9 and using a 30 degree tapered contraction
and a fan inlet velocity pressure of 0.84 "wg [0.21 kPa], the
additional system loss for the inlet transition is -0.31 "wg [-0.08
kPa] and the static pressure at the inlet of the fan is now (-7.0)
"wg + (-0.31) "wg, or -7.31 "wg [-1.74 + (-0.08) = -1.82 kPa].
Knowing the fan inlet velocity pressure, both the revised system static pressue and the fan static pressure can now be calculated to be 7.47 "wg [1.86 kPa].
Using the following equations, the actual system energy
requirement can be calculated using either fan static pressure
or fan total pressure. This example uses fan total pressure, as
total pressure accounts for the total energy from the static and
velocity pressures. To calculate the system energy using fan
static pressure, use fan static pressure and fan static efficiency
instead of fan total pressure and fan total efficiency. Note that
from Equations 7.1, 7.2 and 7.3, FTP = FSP + VPout.
MTR-PWRoutput-std = (MTR-PWRoutput-act) ÷ df
MTR-PWRoutput-act =
MTR-PWRoutput-std = (17.0) ÷ 0.64 = 26.6 hp
[IP]
MTR-PWRoutput-std = (12.8) ÷ 0.64 = 20.0 kW
[SI]
(Flow Rate H Fan Total Pressure) H (1 + Fdrv)
(6343 H Fan Total Efficiency)
= (10000 H (7.47 + 0.84)) H (1 + 0.045) = 18.5 hp
(6343 H 0.74)
[IP]
= (4.7 H (1.86 + 0.21)) H (1 + 0.06) = 13.9 kW
(1.0 H 0.74)
[SI]
and
MTR-PWRoutput-std = (MTR-PWRout-act) ÷ df
MTR-PWRoutput-std = (18.5) ÷ 0.64 = 28.9 hp
[IP]
MTR-PWRoutput-std = (13.9) ÷ 0.64 = 21.7 kW
[SI]
7-46
Industrial Ventilation
Note: When horsepower values published by fan manufacturers do not include the drive loss, the drive loss must be
included in the motor power output calculation to properly size
the motor and determine the motor input power (See Figure 715).
Summary: Comparing the approximated 26.6 hp [20.0 kW]
power requirement in Example 7-6A to the actual 28.9 hp [21.7
kW] power requirement in Example 7-6B, the deviation
between approximate and actual system power is about 8.6%.
System static pressure (SSP) is useful for the designer to calculate the approximate system energy requirement and specify
the fan. Once the fan is selected, the designer should revisit the
system design to ensure that the inlet ducting is properly sized
to the fan inlet, account for any additional losses, and calculate
the final system energy requirement using fan static or total
pressure and fan static or total efficiency. Since system flow
rate and pressure are the variables over which the designer
has the greatest degree of control, a properly calculated system design will minimize the system energy and result in the
most efficient system.
7.5
FAN AND SYSTEM CONTROL
There are a variety of means to control fans and systems
based on the type and range of control required. Restrictive
devices such as outlet dampers increase system resistance to
flow so that the slope of the system curve increases and the
system point of operation moves up and to the left. In these
cases, the fan curve remains fixed while the system curve is
altered by the control method. Other devices such as variable
pitch (axial) fan blades, inlet guide vanes and motor variable
frequency drive controllers change the shape or position of the
fan curve so that the fan curve intersects the system curve at a
different point of operation. In these cases, the fan curve is
altered by the control method and the system curve remains
unchanged. Often, it is desirable to alter both the fan curve and
the system curve for optimum performance, in which case a
combination of control methods would be used.
7.5.1 Flow Control Methods. The most common flow control methods are outlet dampers, variable inlet vane dampers,
variable pitch blades (axial fans) and variable frequency drive
(VFD) motor controllers.
DAMPERS
Outlet dampers are installed at the fan outlet. Because they
are in the air stream, they are subject to material build-up or
wear due to abrasion or erosion from air stream conditions and
may not be acceptable for some applications. Two types of
outlet dampers are available:
1. Parallel Blade Outlet Dampers add system resistance
when partially or fully closed (Figure 7-27). The blades
move in parallel to each other similar to that of a window blind. This damper is the lowest cost but is also
limited in performance as the control arm movement is
not proportional to the degree of control. Its best use is
when only a small degree of control is required
(between 70–100%) or when full open or full closed
operation is needed (such as for cold starts).
2. Opposed Blade Outlet Dampers also add system resistance but over a wider range than parallel blade
dampers. Each blade moves in an opposite direction
(opposing) to the next blade, which allows a more linear relationship between the control arm movement
and the degree of control. This damper is higher cost
but well-suited for applications requiring a broader
range of control or a more even airflow distribution
exiting the damper (Figure 7-27).
3. Variable Inlet Vane and Parallel Blade Inlet Box
Dampers mount at the inlet to the fan or inlet box to
pre-spin the air into the fan wheel. Due to the pre-spin
rotation created by the damper blade position, these
dampers change the shape (pitch) of the fan curve,
reducing fan capacity and operating horsepower. This
allows the fan to operate at high turn-downs without
entering the unstable part of the fan curve (Figure 727). This is helpful with backward-inclined fans, where
surging can take place at as much as 50% of the fan’s
full volume. Because of the power savings, these types
of inlet dampers should be considered for clean air
streams when the fan will operate for long periods at
reduced capacities. However, at inlet damper turndown positions of about 30° open and less, the damper
vanes begin to rotate perpendicular to the air stream
and the inlet damper begins to act as a restriction and
no longer creates a pre-spin vortex. This increase in
static pressure can cause the fan to operate at a point of
instability left of peak on the fan curve as the inlet
damper now increases system resistance (similar to an
outlet damper).
Control dampers can have leakage rates of 10–15% or more
when in the fully closed position. If lower or zero leakage rates
are required, consult the damper supplier for special construction for low or zero leakage options.
VARIABLE PITCH BLADES – AXIAL FANS
Variable pitch blades are available with some axial fan
types. The blades are designed to allow non-standard blade
pitch angles either set by the factory or designed for manual or
automatic changes in the field. Reducing the angle of attack
between the incoming airflow and the blade lowers both the
flow rate and the power demand on the motor. During start-up,
the fan blades can be set to a reduced angle of attack, lowering
the power required to start the fan.
Fans
7-47
FIGURE 7-27. Typical backward-inclined fan curves with volume controls
VARIABLE FREQUENCY DRIVE
A variable frequency drive (VFD) is used to change the fan
speed to create additional fan curves for each change in speed
(Figure 7-27). Because the shape of the fan curves are the
same, reducing the fan speed will mean that flow, pressure and
power will be reduced according to the Fan Affinity Laws.
However, since the VFD only repositions the fan curve to a
higher (increased speed) or lower (reduced speed) curve from
the original fan curve, it is not uncommon to use a damper
along with the VFD in order to control both the fan and system
curves. Common uses for a VFD include reducing motor
power demand during start up or maintaining constant flow by
using a static pressure sensor to control the fan speed in
response to filter media loading in air filtration applications.
The VFD controller is connected between the electric power
source and the motor. Functionally, it varies the voltage and
frequency of the power input to the motor with the motor
speed varying with frequency. For a typical system with fixed
physical characteristics, the attainable points of operation will
fall on the system curve. Figure 7-26 shows points A1 and A2
on a system curve. These two points of operation can be
attained with a VFD by adjusting it for speeds of RPM1 and
RPM2. This will result in fan curve FAN-1 or FAN-2, respectively. If the original design for a system with a pressure loss
across the filter is A2 and the system with a lower pressure loss
is shown as B1-B2, then the VFD can be used to reduce speed,
pressure and power and still maintain a constant flow rate at
Point B1.
VFDs do have disadvantages, including low speed limitations and line noise. Most AC motors are designed to operate
at their nameplate speeds. If a VFD is used to run a motor well
below its nominal speed (50% or less), the motor’s efficiency
may be reduced and (heat) losses can increase. This can
increase motor or VFD heating and may cause damage if the
motor and drive are not designed for these thermal loads. Line
noise can often be addressed with filters and if it is a concern,
consult the VFD supplier.
VFDs can also cause harmonic distortion in the electrical
input lines from the power source. This may affect other electrical equipment on the same power system. This distortion
can be reduced with the addition of isolation transformers or
line inductors or filters. Using isolated motor bearings and/or
motor shaft grounding systems will provide added motor protection. To properly apply a VFD, the supplier needs to know
its intended usage, the building’s power supply, and other electrical equipment in use.
VFDs are commonly used for either Constant Torque or
Variable Torque loads. Examples of constant torque loads are
rotary valves and screw conveyors, while examples of variable
torque loads are fans and pumps. VFDs are normally operated
7-48
Industrial Ventilation
in either a linear volt to hertz ratio common for constant torque
loads or in a squared volt to hertz ratio common for variable
torque loads. Between 0–60 Hz, in a linear volt to hertz ratio
the motor power output varies linearly with the change in
motor speed, while in a squared volt to hertz ratio, the motor
power output varies with the square of the change in motor
speed. Since fans have a power to speed ratio in which the fan
power varies with the cube of the change in fan speed, VFDs
used for fan applications are normally operated in the square
volt to hertz ratio, as this most closely approximates the fan
power to speed ratio and provides the highest operating efficiency. However, there are cases in which operating a fan with
a VFD in a squared volt to hertz ratio results in a higher capital
cost and in these cases, it may be more economical to operate
the VFD in a linear volt to hertz ratio. Since VFDs offer many
different programming options, the VFD supplier should be
consulted for optimal performance.
It is becoming increasingly popular to use VFDs for operation at frequencies greater than 60 Hz. Between 60–90 Hz,
normally power is constant and torque decreases inversely
with the increase in speed (Figures 7-28 and 7-29). Operating
in this range is suitable as long as the torque demand from the
fan is less than the torque output of the motor. The VFD, motor
and fan supplier should always be consulted for these applications. See Figures 7-28 and 7-29 for a comparison of constant
versus variable torque load profiles and Figure 7-30 for a comparison of power savings between outlet dampers, inlet
dampers and VFDs. See Example Problems 7-7 and 7-8 for
use of different flow control methods.
EXAMPLE PROBLEM 7-7 (Flow Control Methods for
System With Flow Turndown at Constant Pressure)
(Figures 7-31 and 7-32) (IP Units Only)
A system is designed to operate using a single fan connected to two (2) parallel main inlet ducts with points of operation at
19000 acfm at 18 "wg SSP and 8000 acfm at 18 "wg SSP. All
measurable static pressure is on the fan inlet and when corrected for duct pressure using a density factor of 0.956, the density
at the fan inlet is 0.0717 lbm/ft3.
To meet the required conditions, the designer can either
select the fan to operate at constant speed and use inlet vane
and opposed blade outlet dampers, or use a variable speed
drive (VFD) and an inlet vane damper.
Case 1: The fan is selected for constant speed with flow control using inlet vane and opposed blade outlet dampers. From
Figure 7-31, the system point of operation of 19000 acfm at
18 "wg SSP intersects the fan curve at 1720 rpm and 67.8
hp. However, when the inlet vane damper is turned down to
its maximum recommended position, the system design point
of operation of 8000 acfm at 18 "wg SSP does not intersect
FIGURE 7-28. Constant torque
the dampened fan curve. In order to intersect the fan and system curves, the opposed blade outlet damper is closed to
increase the system resistance until the system static pressure reaches 21.5 "wg SSP, allowing the system to operate
on the fan curve at a new point of operation of 8000 acfm at
21.5 "wg SSP and 38.0 hp.
Case 2: The fan is selected for variable speed drive with flow
control using a VFD and an inlet vane damper. From Figure
7-32, the system point of operation of 19000 acfm at 18 "wg
SSP intersects the fan curve at 1720 rpm and 67.8 hp. In
order to intersect the fan and system curves at a point that is
“right of peak” on the fan curve, the fan speed is turned down
from 1720 rpm to 1593 rpm and the inlet vane damper is also
dampened to 30 degrees. This allows the system to operate
right of peak on the fan curve and at the design point of 8000
acfm at 18 "wg SSP and 31.6 hp.
For the 19000 acfm condition:
Case 1 and Case 2 are identical and operate at the system
design point of operation of 19000 acfm at 18 "wg SSP.
For the 8000 acfm condition:
Case 1 changes the system design point of operation from
8000 acfm at 18 "wg SSP to 8000 acfm at 21.5 "wg SSP by
using the outlet damper to increase system resistance until
the system curve intersects the fan curve at 8000 acfm and
21.5 "wg SSP.
By using dampers and no VFD, Case 1 has a lower capital
cost, but requires operating at 21.5 "wg SSP and 38.0 hp,
resulting in a higher energy cost and axial loading on the fan.
Case 2 operates at the system design point of operation of
Fans
7-49
FIGURE 7-29. Variable torque
FIGURE 7-30. Controls and power comparison
8000 acfm at 18 "wg SSP by using the VFD to drop (lower)
the fan curve and then using the inlet vane damper to intersect the system curve at the desired point of operation and at
a position that is right of peak on the fan curve.
By using the inlet vane damper and a VFD, Case 2 has a
higher capital cost, but allows operation at the system design
point of operation of 8000 acfm and 18 "wg SSP at 31.6 hp,
resulting in a lower energy cost and axial loading on the fan.
The selection between Case 1 and Case 2 flow control methods should normally be based on the frequency of operation
at the 8000 acfm flow requirement (often or occasional) and
the overall system energy cost.
varies linearly with the change in motor speed from 0–60 Hz,
so the motor power output of the 60 hp motor is
MTR-PWRoutput =
[7.17a]
(MTR-RPMnew ÷ MTR-RPMnameplate) H MTR
PWRnameplate
MTR-PWRoutput = (1629 ÷ 1770) H (60)
MTR-PWRoutput = 55.2 hp, which is sufficient to drive the
fan.
Squared Volt to Hz Output Ratio (Variable Torque):
EXAMPLE PROBLEM 7-8 (Motor Power Determination
for Direct Drive Fan With VFD) (Figure 7-33) (IP Units
Only)
A direct drive fan is selected for 27500 acfm @ 10 "wg FSP,
1629 rpm, 52.6 hp, sea level and 0.0728 lbm/ft3 inlet density
and df = 0.97. The fan is backward inclined with a non-overloading horsepower curve peaking at 53 hp.
Determine the motor horsepower required to drive the fan for
VFD operation in both linear (constant torque) and squared
(variable torque) volt to hz output ratios. Since the fan horsepower curve peaks at 53 hp, assume a 60 hp, 1800 rpm motor
having a synchronous speed of 1770 rpm.
Linear Volt to Hz Output Ratio (Constant Torque):
In a linear volt to hz output ratio, the motor power output
In a squared volt to hz output ratio, the motor power output
varies with the square of the change in motor speed from 0–60
Hz, so the motor power output of the 60 hp motor is
MTR-PWRoutput =
[7.17b]
(MTR-RPMnew ÷ MTR-RPMnameplate)2 H
MTR-PWRnameplate
MTR-PWRoutput = (1629 ÷ 1770)2 H (60)
MTR-PWRoutput = 50.8 hp, which is not sufficient to drive
the fan and a 75 hp motor and VFD is required.
7-50
Industrial Ventilation
FIGURE 7-31. Flow control – constant speed
FIGURE 7-32. Flow control – variable speed
Fans
7-51
FIGURE 7-33. Direct drive fan with VFD control
7.5.2 Fans Operating in Series or Parallel. Sometimes it
is necessary to install two or more fans in a system to provide
a higher pressure or flow rate than can be achieved with a single fan. Two or more fans in series are often used in systems
having high static pressure requirements or in systems having
a process component within the system that has to operate at
or near atmospheric pressure. In environments where ventilation is critical to safety or operations, redundant off-line fans
in parallel may be a safety requirement. When fans are
installed in parallel, all fans are selected for the same static
pressure but may have the same or differing flow rates. In this
case, the system flow rate is additive of each fan, but the static
pressures are not. When fans are installed in series, each fan is
selected for the same flow rate (corrected for inlet densities)
but the fan pressures may vary from fan to fan. In this case, the
static pressure is additive of each fan but the flow rates are not.
For proper application, the fan manufacturer should be consulted for fans operating in parallel or in series (see Figures 734 and 7-35).
2) If axial fans or inline fans are being used, select each
fan for the required (total) flow rate and a proportional
share of the pressure requirements (based on the number of fans used and their location).
3) If centrifugal fans in series are being used, select each
fan for the required (total) flow rate and a proportional
share of the pressure requirements (based on the number of fans used and their location), plus an allowance
for interconnecting ductwork losses. Note that the
above selection process is approximate in that the individual performance of each fan is not the same. All fans
will handle the same mass flow of air but not the same
volumetric flow rate. This is the result of differences in
the air density at the inlet of each fan from differences
in the absolute pressures and air stream temperatures at
the inlet of each fan (such as from heat of compression)
as the air moves through each fan. A rule of thumb for
heat of compression across a centrifugal fan is 1 F temperature rise for each 2 "wg static pressure.
FANS OPERATING IN SERIES
When operating in series, identical fans or fans having the
same wheel types should be used having the same volumepressure curve shapes in order to ensure that all fans operate at
a stable point on the fan curve. The fan manufacturer should
be consulted for selecting fans operating in series. Two fans in
series as shown in Figure 7-34 are selected as follows:
1) Once the system static pressure (SSP) is calculated,
then calculate the Fan Static or Total Pressure (FSP,
FTP) for determining fan selection.
FANS OPERATING IN PARALLEL
4) The required operating range of the system may necessitate two or more fans in parallel instead of one large
fan controlled over a wide operating range. This might
occur when one fan is too large and will not fit into the
desired space or if multiple, lower horsepower fans are
more desirable than a single, larger horsepower fan
(such as in fan arrays).
5) The selection process for fans in parallel requires that
7-52
Industrial Ventilation
Fans
7-53
7-54
Industrial Ventilation
each fan be selected for the same pressure with the flow
rate being the total flow divided by the number of fans
(Figure 7-35).
6) Care must be taken when selecting fans in parallel to
ensure that the system resistance remains on a stable
portion of the fan curves at all times. This is particularly true when the fans have a surge area or dip in the fan
curve (often referred to as “left of peak”). It is important to ensure that the operating point with all fans running is no higher than the lowest pressure in the dip of
the curve. This minimizes the possibility of fan surging, flow reversal, unstable motor loading, or fan system effects.
7) Fans in parallel should have some form of isolation or
control to prevent an off-line fan from rotating backwards due to air flowing back through the off-line fan
(pin-wheeling or wind-milling). A mechanical backstop clutch or a zero leakage isolation valve is recommended to protect against pin-wheeling. In many cases,
a VFD can serve this purpose. Note that a back-stop
clutch or VFD will only provide mechanical protection
against pin-wheeling, while a zero leakage isolation
valve will provide both mechanical protection and prevent reverse migration of air across the fan(s) and into
the system.
7.6
FAN SYSTEM EFFECTS
Fan System Effect is the estimated loss in fan performance
due to non-uniform flow entering or exiting the fan. Fan published ratings data are based on laboratory test conditions having uniform flow entering and exiting the fan. Non-uniform
flow at the fan inlet or outlet often occurs due to the effect of
elbows, dampers, restrictive duct sizing or shape, insufficient
outlet ducting, etc. These conditions can result in severely
degraded fan performance, increased noise and a lower than
expected fan performance.
in pressure between Points 1 and 2. This additional pressure
loss has to be added to the system static pressure calculations
to achieve the design flow rate, and the fan is then selected for
operation at Point 2. Note that because fan system effects are
velocity related, the difference between Points 1 and 2 is
greater than between Points 3 and 4.
7.6.2 System Effect Values. Fan system effect values are
expressed as System Effect Losses (SEL) and are the static
pressure loss in inches of water gauge ("wg) for each system
component having a system effect. The system effect loss
(SEL) can either be derived directly as a static pressure loss
(SEL, "wg) as shown in Figure 7-37 or calculated as shown in
Figure 7-38(7.4) by multiplying a System Effect Factor (SEF) by
the velocity pressure of the component in which the system
effect occurs. The system designer can either enter the system
effect loss (SEL, "wg) in the calculation worksheet as “other
loss” or enter the system effect factor (SEF) in the calculation
worksheet as “special fitting loss factor,” in which case the
actual loss is included in the worksheet calculation. When
more than one system effect is present, the individual system
effect losses (SEL, "wg) are determined and then summed for
entry into the calculation worksheet as “other losses, "wg”, or,
as long as the velocity pressure is common to each system
effect, the individual system effect factors are determined and
then summed for entry into the calculation worksheet as “special fitting loss factor.” Note that when entering the system
effect loss (SEL, "wg) in the calculation worksheet as an
“other loss,” it should be corrected from standard conditions to
the actual conditions by the density factor used in the calculation worksheet using Equation 7.18.
The fan installation must consider the design of the inlet and
outlet ducting to minimize system effects. Many fan system
effects can be avoided or minimized by designing the system
so that the fan inlet and outlet ducting is of sufficient straight
length and diameter entering and exiting the fan. If system
effect conditions cannot be avoided, the system designer must
calculate and add the system effect for each condition to the
System Static Pressure calculations prior to selecting the fan.
7.6.1 Impact on System Performance. Figure 7-36 illustrates deficient fan system performance due to unaccounted for
fan system effects. The system pressure losses have been
determined accurately and a fan is selected for operation at
Point 1. Because no allowance has been made for fan system
effects, the point of intersection between the fan and the actual
system curve is at Point 4. Unless corrected, the actual flow
rate will be deficient by the difference in flow between Points
1 and 4. The fan system effect loss is shown by the difference
FIGURE 7-36. Fan System Effect (FSE)
Fans
SYSTEM EFFECT LOSS, "wg (SEL)
Figure 7-37 [IP, SI] shows a series of System Effect Curves.
By entering the chart at the appropriate air velocity along the
x-axis, it is possible to read across from any curve to the y-axis
and determine the system effect loss (SEL) in "wg for a specific condition. Note that the system effect curves are plotted
for standard air density using 0.075 lbm/ft3 [1.204 kg/m3]. For
non-standard air, the system effect loss (SEL) is proportional
to density and can be calculated by:
SELact = (SELstd) (df) or
SELact = (SELstd) (ρact/ρstd)
[7.18]
where:
SELact = system effect loss, "wg [Pa] at actual conditions
SELstd = system effect loss, "wg [Pa] at standard
conditions
df = air stream density factor
ρstd = standard air stream density, 0.075 lbm/ft3
[1.204 kg/m3]
ρact = actual air stream density, lbm/ft3 [kg/m3]
USING SYSTEM EFFECT LOSS FACTORS (SEF) TO
DETERMINE SYSTEM EFFECT LOSS (SEL, "wg)
The system effect loss (SEL, "wg) can also be determined
using the system effect factors designated for the system effect
curves shown in Figure 7-38. The system effect loss is determined by multiplying the appropriate system effect factor
(Fsys) in Figure 7-38 by the velocity pressure (VP) at the
appropriate system component. When the system effect factor
(SEF) is entered into the calculation worksheet as “special fitting loss factor,” the multiplication is a part of the worksheet
calculation.
7.6.3 Fan Inlet System Effects. Some of the more common
conditions causing system effects at the fan inlet include:
FAN INLET ELBOWS
Non-uniform flow into a fan inlet is a common cause of
deficient fan performance. Any elbow located close to the fan
inlet will not allow uniform entry of the air into the fan wheel.
The result is a performance loss by the fan. System effect
curves and guidelines for inlet duct elbows are given in
Figures 7-39, 7-40, 7-41, 7-42 and 7-43.
FAN INLET BOXES
In an attempt to reduce system effects due to elbows at the
inlets of centrifugal fans, fan manufacturers design and provide special fittings called inlet boxes. Most fan manufacturers
recommend an additional 0.75 VP loss due to system effects
but even the best designed inlet box may have a loss in excess
7-55
of 1 VP. The actual system effect for a fan inlet box should be
obtained from the fan manufacturer. In the absence of fan
manufacturer’s data, a well designed inlet box should approximate system effect curves S or T in Figures 7-37 and 7-38.
When the inlet box is equipped with a damper, if the pressure
loss is not included in the fan selection software by the fan
manufacturer, the pressure loss across the damper should be
included in the system static pressure calculations.
FAN INLET SPIN
Another cause of reduced performance is when the air on
the inlet side of the fan has a vortex or spin as the air enters the
fan as shown in Figure 7-40. The preferred inlet condition
allows air to enter the fan without spin in either direction. A
spin in the same direction as the fan rotation (pre-rotation)
reduces the load on the fan and shifts the fan performance
curve down and to the right by an amount dependent on the
intensity of the vortex. The effect is similar to the change in the
fan performance curve by a variable inlet vane damper at the
fan inlet in which the vanes induce a controlled spin in the
direction of fan rotation, reducing the flow rate, pressure and
power. A counter-rotating spin rotates the air in the opposite
direction of the fan. This spin results in an increase in fan performance, noise and power.
A vortex or spin of the air entering the fan inlet can also be
created by non-uniform flow conditions from an upstream system component, such as a cyclone or back to back offset
elbows. Since the causes of inlet spin are both numerous and
variable, its system effects cannot be recorded. Where a vortex
or inlet spin cannot be avoided, the use of turning vanes, splitter sheets, or egg crate straighteners will reduce the degree of
the fan system effect. See AMCA Publication 201.
FAN INLET VANE DAMPER OR INLET BOX DAMPER
When a variable inlet vane damper or an inlet box with an
inlet damper is supplied by the fan manufacturer, a system
effect has to be included either by the system designer or the
fan manufacturer. Figure 7-44 shows the system effect curves
for typical inlet vane dampers. Consult the fan manufacturer
for the effect of an inlet box damper.
When space limitations prevent optimum fan inlet conditions, better flow conditions can be achieved by using turning
vanes in the inlet elbow (Figures 7-40 and 7-42). The turning
vanes should be rigid and run from throat to throat of the
elbow (inlet to outlet). Normally, there are at least two (2)
vanes, with the vanes positioned based on equal flow areas.
The pressure loss of these elbows with turning vanes should be
added to the system pressure losses.
OBSTRUCTED FAN INLET
A fan system effect occurs when there is an obstruction to
flow at the fan inlet. The most common inlet obstruction is
when the fan inlet duct diameter is smaller or larger than the
7-56
Industrial Ventilation
Fans
7-57
7-58
Industrial Ventilation
Fans
7-59
7-60
Industrial Ventilation
Fans
7-61
7-62
Industrial Ventilation
Fans
7-63
7-64
Industrial Ventilation
Fans
diameter of the fan inlet. The system effects for obstructed fan
inlets are shown in Figure 7-45. The system effect curves
shown are based on the ratio of the area of the inlet duct to the
area of the fan inlet. For inlet ducts having areas greater than
the fan inlet area, treat the system effect as a sharp edge orifice
entry loss and use a 1.78 SEF to calculate the system effect
loss (SEL).
7.6.4 Fan Outlet System Effects. Figures 7-46 and 7-47
show changes in velocity profiles of the air at various distances
from centrifugal and axial flow fan outlets. The velocity profile at the outlet of any fan is not uniform. Since air has both
mass and inertia, it concentrates along the outer portion of the
fan housing. This results in higher velocities above the blast
area and lower velocities below the blast area for centrifugal
fans, and higher velocities along the housing walls and lower
velocities in the center for axial fans. A straight section of duct
is required to establish a more uniform air velocity profile. For
axial and centrifugal fans, the air will reach uniform flow at a
100% effective duct length, defined as 2.5 equivalent duct
diameters for velocities less than 2500 fpm [13 m/s]. For
velocities greater than 2500 fpm [13 m/s], add 1 duct diameter
for each additional 1000 fpm [5 m/s]. For example, a fan discharge duct with a velocity of 3500 fpm would require a length
of 3.5 equivalent duct diameters to achieve an effective 100%
duct length. Note that for centrifugal fans, 90% of the system
effect loss is recovered with only 50% effective duct length. As
such, for large or limited space discharge ducts, a 50% effective duct length can be used with minimal energy loss. For
axial fans, system effect values are under review and not listed
for effective duct lengths greater than 25%.
7-65
SELact = (SELstd) (df) or SELact = (SELstd) (ρact/ρstd)
If the system effect factor (SEF) is determined from Figure
7-38 and included in the system calculation worksheet as part
of the system static pressure (SSP) calculation, then no further
corrections are necessary.
If the system effect loss (SEL) is not included in the system
calculation worksheet, then it can either be added to the system
static pressure or fan static or total pressure values as an additional static loss as follows:
If using loss values directly from Figure 7-37 or if using
Figure 7-38 and multiplying the SEF by the velocity pressure
at standard conditions, the SEL has to be corrected for the density factor:
SEL = [(|SEL1| + |SEL2| + • • • + |SELn|)] (df)
[7.19a]
If using Figure 7-38 and multiplying the SEF by the velocity pressure at actual conditions, the SEL is determined by:
SEL = [(|SEL1| + |SEL2| + • • • + |SELn|)]
[7.19b]
Since fan system effects occur at or in close proximity to the
fan inlet, the density factor used for the fan system effect calculation is the density factor at the fan inlet. In special cases
where the designer is correcting for compressibility on the fan
outlet, the density factors for the fan inlet and outlet will not be
the same and the appropriate density factor should be used.
FAN OUTLET ELBOWS
Since the velocity profile at the outlet of the fan is not uniform, an elbow located at or near the fan outlet will degrade
fan performance. Elbows at the fan outlet should be placed
downstream of the 100% effective duct length to avoid system
effects. When outlet elbows cannot be avoided, they should
follow the same direction as the wheel rotation. Elbows
installed at less than the 100% effective outlet duct length and
turning in the opposite direction of the wheel may result in
severe fan performance loss and increased noise – a condition
referred to as “breaking the back of the fan.” However, when
the elbow is installed at distances equal to or greater than
100% effective duct length, the only loss is that of the elbow
itself. Figures 7-48 and 7-49 show system effect curves for
centrifugal and vaneaxial fans. Tubeaxial fans with two and
four piece elbows have a negligible SEF and are not shown.
7.6.5 Calculating Fan System Effects. Fan system effect
losses (SEL, "wg) are derived either directly from Figure 7-37
for standard conditions or by multiplying the system effect factor (SEF) from Figure 7-38 by the velocity pressure of the fitting in which the system effect occurs. If the system effect loss
(SEL, "wg) is calculated for standard conditions, then the loss
has to be corrected for actual conditions using Equation 7-18:
EXAMPLE PROBLEM 7-9 (Calculation of System Effect
Factors) (Figure 7-50) (IP Units)
In Figure 7-50, a fan has a 13" diameter (0.92 ft2), 4-piece,
90° round inlet elbow with a 2.0 centerline radius at the fan
inlet. The fan does not have an outlet duct. The fan inlet and
outlet area is 0.92 ft2 and the fan blast area is 0.74 ft2. The
required flow rate is 4000 acfm and the Fan Static Pressure is
8 "wg at standard conditions (df = 1.0). Selecting a fan without
the fan system effect and using Table 7-1, results in a fan
speed of 1,862 rpm and power consumption of 9.5 hp.
Corrections for fan system effects are:
Fan Inlet System Effect:
Using Figures 7-38 and 7-39, the system effect curve is found
to be between R-S, so a 1.0 SEF is selected. Since V = Q/A =
4348 fpm and
VP = (V/4005)2 (df) = (4348/4005)2 (1) = 1.18 "wg
System Effect Loss, Inlet =
(SEF) (VPinlet) = (1.0) (1.18 "wg) = -1.18 "wg
7-66
Industrial Ventilation
Fans
7-67
7-68
Industrial Ventilation
Fans
7-69
7-70
Industrial Ventilation
Fans
7-71
Manufacturers Association (NEMA). These standards apply to
motor design parameters such as horsepower ratings, dimensions, enclosures, power requirements, and insulation.
7.7.1 Motor Selection Criteria.
MOTOR SUPPLY POWER
Voltage (V): The supply voltage must be known for motor
and controls selection. Three phase power is generally either
230 or 460 volts in the United States and 575 volts in Canada.
Single phase power is usually either 115 or 230 volts.
FIGURE 7-50. Fan inlet elbow (Example Problem 7-9)
Fan Outlet System Effect:
Using Figures 7-46 and 7-38 and a fan blast to outlet area ratio
of 0.74/0.92 = 0.8, the system effect curve is between T-U, so
a 0.45 SEF is selected.
System Effect Loss, Outlet =
(SEF) (VPoutlet) = (0.45) (1.18 "wg) = 0.53 "wg
The System Effect Loss is then
SEL = [|SEL1| + |SEL2|] = [1.18 + 0.53] = 1.71 "wg
The Fan Static Pressure is then corrected to include the fan
system effect loss:
FSPcorr = FSP + SEL
FSPcorr = 8 "wg + 1.71 "wg = 9.71 "wg
The fan would be re-selected for 4000 acfm @ 9.71 "wg FSP
and a 1.0 density factor.
7.7
FAN MOTORS
Most fans are driven by electric motors. Selecting the correct motor requires information on the electrical power available, the fan power demand, fan speed, how the fan is driven
(belt or direct drive), and the ambient conditions where the fan
and motor are located.
In North America, electric motor manufacturers comply
with standards established by the National Electric
Current (Amps): The most common form of supply power
is alternating current (AC). Direct current (DC) can be used in
special-purpose cases where variable speed is needed and the
torque load is low, but this is changing with the availability and
lower cost of variable frequency AC drives.
Phase (3 or 1): Power is supplied by either a 3-wire, three
phase or a 2-wire, single phase system. Three phase is typical
for industrial applications and delivers the current between
three wires instead of two, reducing the load on each wire. It
is also more economical as it requires smaller lead wires due
to the lower current load for each wire. Single phase is commonly used for fractional horsepower motors.
Frequency (Hertz, Hz): Frequency, in cycles per second, is
either 50 or 60 Hz, depending on the location. Standard frequency for the United States and Canada is 60 Hz, while many
other countries use 50 Hz.
MOTOR CONSTRUCTION
Motor Voltage (V): The voltage(s) listed on the motor nameplate is the design voltage(s) of the motor. Motors are typically
capable of operating within ± 10% of the nameplate voltage.
Full Load Amps (FLA): The motor full load amps listed on
the motor nameplate is the current the motor will draw when
running at nameplate power. Measuring the actual motor amps
and comparing it to the FLA is sometimes a quick way to estimate the actual motor load.
Power Rating (HP, W, kW): The motor power output must
be greater than the power demand from the fan and drive system. The power rating listed on the motor nameplate is the
maximum power the motor will produce at its rated speed.
Motors in North America are rated in horsepower (hp); in
much of the rest of the world motors are rated in watts or kilowatts (W, kW).
Torque = (MTR-PWRoutput H 5250) ÷ RPM
where:
Torque = lb-ft
MTR-PWRoutput = Motor output power, hp (W, kW)
5250 = conversion constant
RPM = motor speed
7-72
Industrial Ventilation
MTR-PWRoutput-act (watts) =
[7.20a]
(Volts H Amps H CF H Efficiency H PF)
MTR-PWRoutput-act (horsepower) =
[7.20b]
(Volts H Amps H CF H Efficiency H PF) ÷ 746
MTR-PWRoutput-std = (MTR-PWRout-act) ÷ df
(Equation 7.16)
Power Factor (PF): The power factor for three phase power
is a measure of phase difference between voltage and current in
the motor circuit. The power factor value is always less than 1.0.
where:
MTR-PWRoutput = motor output power, hp (W, kW)
Volts = avg supply voltage to motor
Amps = avg amp draw at motor
CF = 1.732 for 3 phase motors (√3)
1.0 for single phase motors
Efficiency = motor rated efficiency
PF = motor power factor
746 = unit conversion from watts to
horsepower
Motor Frame: NEMA sets industry standards for certain
motor dimensions and designates them by frame size. Motors
with common frame sizes have the same shaft diameter, centerline height, and mounting dimensions. To determine the
shaft centerline height of a NEMA motor, divide the first two
(2) digits of the motor frame size by four (4) to calculate the
motor shaft centerline height in inches. Example: 182T frame
motor shaft centerline height = 18/4 = 4.5 inches.
Motor Service Factor (SF): The motor service factor is a
multiplier applied to the motor nameplate power for intermittent service above the nameplate power when rated voltage and
frequency is supplied to the motor. Motor service factors range
from 1.0 to 1.25, depending on the application and power with
1.15 being most common. Motors on inverter drives (VFDs)
will normally operate with a 1.0 service factor.
Motor Speed (RPM): AC motor speed is a function of the
line frequency and the number of poles in the motor.
Nx = (120) (Hz)/p
[7.21]
where:
Nx = motor synchronous speed (rpm)
Hz = line frequency, 50 or 60 hertz
p = number of poles
Motors run at speeds slightly below their synchronous
speed. For example, a four pole, 60 Hz motor has a synchronous speed of 1800 rpm but will normally run between
1725 and 1770 rpm. The difference in the motor synchronous
speed and the motor full load speed is called slip. Most motors
will operate with a slip of 5% or less.
Motor Efficiency: Motor efficiency is the ratio of the motor
output power to the electrical input power. Motor efficiency ηm
is not constant but changes with the operating load of the
motor. Motor manufacturers typically supply motor efficiency
values at full, 75%, 50% and 25% load.
Motor Efficiency, ηm =
MTR-PWRoutput-act ÷ MTR-PWRinput-act
Motor Efficiency Standards require motor manufacturers to
certify that their motors meet minimum efficiency values. In
the United States the Energy Independence and Security Act of
2007 (EISA 2007) defines energy efficiency standards for general purpose electric motors and specialty motor designs. The
standards require electric motors to have a nominal full load
efficiency that is equal to or greater than the energy efficiency
defined in National Electrical Manufacturers Association
(NEMA) Standards Publication MG1.
[7.22]
Motor Enclosure: The selection of the motor enclosure
depends on the site conditions. The type of enclosure indicates
the type of protection for the internal motor components from
the ambient environment and the method of motor cooling.
Open Drip-Proof (ODP) motors have openings in the motor
enclosure allowing air movement directly through the motor.
Air drafts into the motor and across the rotor and windings for
cooling. ODP motors should only be used for clean, low moisture, indoor applications.
Totally Enclosed Fan Cooled (TEFC) motors do not have
openings in the motor enclosure, but are not necessarily airtight. An integral cooling fan blows air over the motor enclosure to cool the motor. TEFC motors are used in indoor or outdoor applications where dust and water are present in modest
amounts. They are not water tight or designed to withstand
direct water spray or washing.
Total Enclosed Air Over (TEAO) motors are similar to
TEFC motors except there is no integral cooling fan. These
motors are frequently used on fans where the motor is in the
air stream, providing cooling to the motor.
Totally Enclosed Explosion Proof (TEXP): Explosion Proof
motors are special versions of TEFC motors with design features making them suitable for applications where explosive
dust or gases are present. The enclosure is designed to withstand an explosion inside the motor and contain the flame and
sparks within the motor. There are different NEMA classifications of explosion proof construction, depending on the characteristics of the explosive gas or dust.
Severe Duty Motors are another variation of TEFC motors
having features that make them durable in hostile environments. They have better shaft seals, corrosion resistant paint,
and are available with stainless steel shafts.
Fans
Temperature Ratings: Motors are available in different temperature ratings which are identified by insulation classes. The
most common insulation class is Class B, which is used for
general purpose applications. Class F and H insulation are
used in motors for high ambient temperature applications.
High temperature applications may occur from frequent overloading of the motor, from the use of variable frequency drives
or from ambient conditions greater than 104 F [40 C].
Motor Inertia Load Capacity: In some cases, it is not only
the power requirements that determine the size of motor but
the number of times the motor is started and the motor’s ability
to accelerate the fan to full speed. This is particularly true
when using small motors on large fans. Motors must have an
inertia load capability greater than the inertia of the fan corrected for the drive ratio, as shown in the equation below:
WK2motor > WK2fan (RPMfan ÷ RPMmotor)2 (1.1)
REFERENCES
7.1
Gibson, N.; Lloyd, F.C.; Perry, G.R.: Fire Hazards in
Chemical Plants from Friction Sparks Involving the
Thermite Reaction. Symposium Series No. 25.
Institute of Chemical Engineers London (1968).
7.2
Air Movement and Control Association, Inc.:
Standards Handbook, AMCA Publication 99-16
(2016).
7.3
Air Movement and Control Association, Inc.: Field
Performance Measurement of Fan Systems, AMCA
Publication 203-90 (2011).
7.4
Air Movement and Control Association, Inc.: Fans and
Systems, AMCA Publication 201-02 (R2011).
7.5
Air Movement and Control Association, Inc.: Air
Systems, AMCA Publication 200-95 (R2011).
7.6
Air Movement and Control Association, Inc.: Vaneaxial Fan, Exploded View, AMCA Publication 99-16
(2016).
[7.23]
where:
WK2motor = inertia load by the motor at the shaft
WK2fan = inertia load of the fan
RPMmotor = motor speed
RPMfan = fan speed
1.1 = 10% factor for V-belt drives
If the motor does not have enough inertia capacity, either it
will not be able to start the fan or it will take an excessive
amount of time to bring the fan up to speed. If this occurs, consult the fan or motor manufacturer for acceptable solutions.
7.7.2 Motor Installation. The National Electric Code specifies the requirements for motor installation and wiring. The
sizing of motor lead wires and overload protection must consider any higher than normal amp draw that occurs when a
motor is started and brought up to full speed. For across the
line starting, an in-rush current load of 6 to 10 times the motor
full load nameplate amps is not uncommon (for a few seconds). As a result, motor branch circuits for fans are often
sized differently than other types of branch circuits. There are
also requirements that specify how close motors and disconnects should be located. These are important since they provide personnel protection for servicing the fan and motor.
7-73
ACKNOWLEDGMENTS
Fan drawings depicted in Figures 7-3, 7-4, 7-5, 7-6 and 7-7 are
courtesy of Twin City Fan Company.
Figures 7-28, 7-29, 7-30, 7-31, 7-32 and 7-33 are courtesy of
The New York Blower Company and M&P Air Components,
Inc.
The Affinity Laws are adapted from Fan Engineering, An
Engineer’s Handbook on Fans and Their Applications, Eighth
Edition, by Buffalo Forge Company.
If a fan is belt driven, the motor must be mounted on an
adjustable base. When loosened, this base allows motor movement for aligning the drive and tensioning and replacing the
belts. Figure 7-51 shows the AMCA designated motor positions for centrifugal belt driven fans. For direct drive fans, the
motor mounting base should include horizontal alignment
blocks or a similar device allowing motor adjustment both parallel to and perpendicular to the motor shaft. For all fans,
motors should only be adjusted in the vertical plane using
machined shims.
FIGURE 7-51. Motor locations for belt driven centrifugal fans
Air Cleaning Devices
8-1
Chapter 8
AIR CLEANING DEVICES
NOTE: Equations with notation followed by (IP) are designated for inch-pound system only; equations followed by (SI) are designated for
metric use only. If equation bears neither, then it applies to both systems.
8.1
8.2
8.3
8.4
8.5
8.6
INTRODUCTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-2
8.7 UNIT COLLECTORS . . . . . . . . . . . . . . . . . . . . . . . . . 8-33
SELECTION OF DUST COLLECTION
8.8 DUST COLLECTING EQUIPMENT COST . . . . . . . 8-33
EQUIPMENT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-2
8.8.1 Price versus Capacity . . . . . . . . . . . . . . . . . . . 8-33
8.2.1 Efficiency Required . . . . . . . . . . . . . . . . . . . . . 8-2
8.8.2 Accessories Included . . . . . . . . . . . . . . . . . . . 8-33
8.2.2 Gas Stream Characteristics . . . . . . . . . . . . . . . . 8-3
8.8.3 Installation Cost . . . . . . . . . . . . . . . . . . . . . . . 8-37
8.2.3 Contaminant Characteristics . . . . . . . . . . . . . . . 8-3
8.8.4 Special Construction . . . . . . . . . . . . . . . . . . . . 8-37
8.2.4 Energy Considerations . . . . . . . . . . . . . . . . . . . 8-3
8.9 SELECTION OF DISPOSABLE-TYPE AIR
8.2.5 Dust Discharge and Disposal . . . . . . . . . . . . . . 8-3
FILTRATION EQUIPMENT . . . . . . . . . . . . . . . . . . . . 8-37
DUST COLLECTOR TYPES . . . . . . . . . . . . . . . . . . . . 8-3
8.9.1 Straining . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-37
8.3.1 Electrostatic Precipitators . . . . . . . . . . . . . . . . . 8-3
8.9.2 Impingement . . . . . . . . . . . . . . . . . . . . . . . . . . 8-37
8.3.2 Fabric Collectors . . . . . . . . . . . . . . . . . . . . . . . . 8-8
8.9.3 Interception . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-37
8.3.3 Wet Collectors . . . . . . . . . . . . . . . . . . . . . . . . . 8-18
8.9.4 Diffusion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-37
8.3.4 Dry Centrifugal Collectors . . . . . . . . . . . . . . . 8-21
8.9.5 Electrostatic . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-37
ADDITIONAL AIDS IN DUST COLLECTOR
8.9.6 Disposable Filter Rating . . . . . . . . . . . . . . . . . 8-37
SELECTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-25
8.10 RADIOACTIVE AND HIGH TOXICITY
CONTROL OF MIST, GAS AND VAPOR
OPERATIONS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-39
CONTAMINANTS . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-25
8.11 EXPLOSION VENTING/DEFLAGRATION
GASEOUS CONTAMINANT COLLECTORS . . . . . 8-25
VENTING . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-40
8.6.1 Absorption . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-30
REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-41
8.6.2 Adsorption . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-32
APPENDIX A8 CONVERSION OF POUNDS PER
8.6.3 Incineration/Oxidation . . . . . . . . . . . . . . . . . . 8-32
HOUR (EMISSIONS RATE) TO GRAINS
8.6.4 Biofiltration . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-33
PER DRY STANDARD CUBIC FOOT . . . . . . . . . . . .8-42
8.6.5 Other Gaseous Contaminant Controls . . . . . . .8-33
____________________________________________________________
Figure 8-1
Figure 8-2
Figure 8-3
Figure 8-4
Figure 8-5
Figure 8-6
Figure 8-7
Figure 8-8
Figure 8-9
Figure 8-10
Figure 8-11
Figure 8-12
Table 8-1
Table 8-2
Table 8-3
Dry Type Dust Collectors – Dust Disposal . . 8-4
Figure 8-13
Wet Type Collectors (for Particulate
Dry Type Dust Collectors – Discharge Valves .8-5
Contaminants) . . . . . . . . . . . . . . . . . . . . . . . . 8-24
Dry Type Dust Collectors – Discharge Valves .8-6
Figure 8-14
Dry Type Centrifugal Collectors . . . . . . . . . 8-26
Figure 8-15 (IP, Range of Particle Size and
Electrostatic Precipitator High Power Design
SI)
Collector Efficiencies . . . . . . . . . . . . . . . . . . 8-27
(40,000 to 75,000 Volts) . . . . . . . . . . . . . . . . . 8-7
Figure 8-16
Characteristics of Particles and Particle
Electrostatic Precipitator Low Power Design
Dispersoids . . . . . . . . . . . . . . . . . . . . . . . . . . 8-31
(11,000 to 15,000 Volts) . . . . . . . . . . . . . . . . . . 8-9
Figure 8-17
Unit Collector (Shaker Type Fabric) . . . . . . 8-34
Performance vs. Time Between
Figure 8-18 (IP) Cost Estimates of Dust Collecting
Reconditionings for Fabric Collectors . . . . . 8-12
Equipment . . . . . . . . . . . . . . . . . . . . . . . . . . .8-35
Baghouse Cleaning Mechanisms . . . . . . . . . 8-15
Figure 8-18 (SI) Cost Estimates of Dust Collecting
Fabric Collectors – Pulse Jet Type . . . . . . . . 8-17
Equipment . . . . . . . . . . . . . . . . . . . . . . . . . . .8-36
Interstitial Velocity Calculations . . . . . . . . . 8-19
Figure 8-19
Comparison Between Various Methods of
Dust Containment Booth . . . . . . . . . . . . . . . 8-20
Measuring Air Cleaning Capability . . . . . . . 8-39
Wet Type Collector (for Gaseous
Contaminant) . . . . . . . . . . . . . . . . . . . . . . . . . 8-22
Wet Type Dust Collectors (for Particulate
Contaminants) . . . . . . . . . . . . . . . . . . . . . . . . 8-23
____________________________________________________________
Characteristics of Filter Fabrics . . . . . . . . . . . .8-11
Summary of Fabric Type Collectors and
Their Characteristics . . . . . . . . . . . . . . . . . . . . .8-14
Dust Collector Selection Guide . . . . . . . . . . . .8-28
Table 8-4
Table 8-5
Table 8-6
Comparison of Some Important Dust
Collector Characteristics . . . . . . . . . . . . . . . . .8-38
Media Velocity vs. Fiber Size . . . . . . . . . . . . .8-39
Comparison of Some Important Air Filter
Characteristics . . . . . . . . . . . . . . . . . . . . . . . . . .8-40
8-2
8.1
Industrial Ventilation
INTRODUCTION
Air cleaning devices remove or render harmless contaminants from an air or gas stream. They are available in a wide
range of designs to meet variations in air cleaning requirements. Degree of removal required, typically dictated by governmental standards, quantity and characteristics of the contaminant to be removed, and conditions of the air or gas stream
will all have a bearing on the device selected for any given
application. In addition, fire safety and explosion control
should be considered in all selections (see Section 8.11).
This chapter will give an overview of major contaminant
control devices, whether the contaminant is in solid, liquid
(aerosol) or in a gaseous state. In order to choose the proper
control device, it is of absolute importance to know the chemical constituents, particle or aerosol size distribution and relative concentration of those pollutants. The U.S. Environmental
Protection Agency (USEPA) has accepted methods of determining the constituents of different air streams. Testing done
outside of these sanctioned test methods are likely not to be
accepted as proof of compliance (see www.epa.gov).
For particulate contaminants, air cleaning devices are
divided into two basic groups: air filters and dust collectors.
Air filters are designed to remove low dust concentrations of
the magnitude found in atmospheric air. They are typically
used in ventilation, air-conditioning, and heating systems
where dust concentrations seldom exceed 1.0 grains per thousand cubic feet of air, and are usually well below 0.1 grains
per thousand cubic feet of air [0.23 mg/m3]. (One pound
equals 7,000 grains. A typical atmospheric dust concentration
in an urban area is 87 micrograms per cubic meter or
0.000038 grains per standard cubic foot of air.)
Dust collectors must be capable of handling concentrations
100 to 20,000 times greater than those for which disposable air
filters are designed.
8.2
SELECTION OF DUST COLLECTION EQUIPMENT
Dust collection equipment is available in numerous designs
utilizing many different principles and featuring wide variations in effectiveness, first cost, operating and maintenance
cost, space, arrangement, and materials of construction.
Consultation with the equipment manufacturer is the recommended procedure in selecting a collector for any problem
where extensive previous plant experience on the specific dust
problem is not available.
8.2.1 Efficiency Required. Previously, there was no accepted standard for testing and/or expressing the efficiency of a
dust collector. It was virtually impossible to accurately compare the performance of two collectors by comparing efficiency claims. The filter materials could be, and still are, compared
based on the minimum efficiency reporting value (MERV) rating (see Section 8.9.6). However, ANSI/ASHRAE has recently developed Test Method 199-2016 to provide operational
performance including emissions and total energy consump-
tion under simulated operating conditions. This test method is
intended to allow owners to compare performance of different
design collectors using third party test data. Process applications involving temperature, humidity, and non-standard gas
compositions do not lend themselves to standardized testing.
In these cases, consumers should consider the collector manufacturer’s experience with the same or similar processes.
Currently, ANSI/ASHRAE Method 199-2016 applies only to
fabric collectors. Evaluation of performance for other equipment types such as high voltage electrostatic precipitators, oxidizers, wet collectors, or dry centrifugal collectors, will require
field measurements. The best measure of performance for any
collector is the actual mass emission rate, expressed in terms
such as mg/m3 or grains/ft3, conducted under actual operating
conditions.
When the cleaned air is to be discharged outdoors, the
required degree of collection can depend on facility location,
nature of contaminant (its salvage value and its potential as a
health hazard, public nuisance, or ability to damage property),
and the regulations of governmental agencies. In remote locations, damage to eco-sensitive farms or contribution to air pollution problems of distant cities can influence the need for and
importance of effective collection equipment. Many industries,
originally located away from residential areas, failed to anticipate the construction of residences that frequently develop
around a facility. Such lack of foresight has required installation of air cleaning equipment at greater expense than initially
would have been necessary. Today, the remotely located manufacturer should comply, in most cases, with the same regulations as if it were located in an urban area. With present and
future emphasis on public nuisance, public health, and preservation and improvement of community air quality, management can continue to expect criticism for excessive emissions
of air contaminants whether located in a heavy industry section
of a city or in an area closer to residential zones.
A safe recommendation in equipment selection is to select
the collector that will allow the least possible amount of
contaminant to escape and is reasonable in first cost and
maintenance while meeting all prevailing air pollution regulations. For some applications, even the question of reasonable cost and maintenance should be sacrificed to meet established standards for air pollution control or to prevent damage
to health or property. However, in areas designated as above
the established National Ambient Air Quality Standards
(NAAQS) for a pollutant, for example, multiple control
devices may be required in order to minimize emissions to the
lowest achievable emission rate (LAER) as designated by the
USEPA.
It should be remembered that visibility of an effluent will be
a function of the light reflecting surface area of the escaping
material. Surface area per pound increases inversely as the
square of particle size. This means that the removal of 80% or
more of the dust on a weight basis may still remove only the
coarse particles without altering the stack appearance.
Air Cleaning Devices
8.2.2 Gas Stream Characteristics. The characteristics of
the carrier gas stream can have a significant impact on equipment selection and performance. Temperature of the gas
stream may limit the media choice in fabric collectors. High
temperatures and low gas densities will reduce the collection
efficiencies for centrifugal collectors. Condensation of water
vapor will cause packing and plugging of air or dust passages
in dry collectors. Corrosive chemicals can attack fabric or
metal in dry collectors and when mixed with water in wet collectors can cause extreme damage.
8.2.3 Contaminant Characteristics. The contaminant characteristics will also affect equipment selection. Sticky materials, such as metallic buffing dust impregnated with buffing
compounds, can adhere to collector elements and plug collector passages. Linty materials can adhere to certain types of collector surfaces or elements. Abrasive materials such as mill
scale or silica in moderate to heavy concentrations will cause
rapid wear on dry metal surfaces. Particle size, shape, and density (specific gravity) can reduce collection efficiency of centrifugal collectors. For example, the parachute shape of particles like the bee’s wings from grain are more challenging to
collect because their shape causes them to behave like a much
smaller particle having a low terminal velocity. The equivalent
spherical diameter of these particles is referred to as the aerodynamic particle size. In addition, the combustible nature of
many finely divided materials will require specific collector
designs to assure safe operation.
Contaminants in exhaust systems cover an extreme range in
concentration and particle size. Concentrations can range from
less than 0.1 to more than 10 grains of dust per cubic foot of
air [0.229 g/m3 to 22.9 H 105 g/m3] and in excess of 100 grains
per cubic foot [229 g/m3] for pneumatic conveying systems. In
low pressure conveying systems, the dust ranges from 0.5 to
100 or more microns in size. Deviation from mean size will
also vary with the material (Figure 8-15).
8.2.4 Energy Considerations. The cost and availability of
energy makes essential the careful consideration of the total
energy requirement for each collector type that can achieve the
desired performance. The cost of all energy sources should be
considered when evaluating collector technologies such as fan
power, pump power, compressed air, etc. An electrostatic precipitator, for example, might be a better selection at a significant initial cost penalty because of the energy savings due to
its lower pressure drop.
8.2.5 Dust Discharge and Disposal. Dust removed from
the collector becomes either a solid waste stream, liquid waste
stream, or is re-introduced to the process. Methods of removal
and disposal of collected materials will vary with the material,
plant process, quantity involved, and collector design. Dry collectors can be unloaded continuously or in batches through
dump gates, trickle valves, and rotary airlocks to conveyors or
containers. Dry materials can create a secondary dust problem
if careful thought is not given to dust-free material disposal.
See Figures 8-1, 8-2, and 8-3 for some typical discharge
8-3
arrangements and valves.
Material should never be stored in the collector hopper
unless it was specifically designed for this purpose. Selection
of rotary valves should consider the material characteristics for
proper selection. Rotor speeds should not be selected such that
they create a “fan effect” in the discharge of the hopper. The
“fan effect” can reduce collection efficiency of centrifugal collectors and reduce airlock capacity. Selection speeds of 20
RPM or less for fabric collectors and 15 RPM or less for centrifugal collectors will usually avoid the “fan effect” and ensure
the rotary valve is able to remove the material as designed.
Wet collectors should have a continual ejection of collected
material unless the recycle tank is specifically designed to separate the solids from the scrubbing water. Secondary dust problems are eliminated although disposal of wet sludge and treatment of liquid slurry can be a material handling problem. Solids
or dissolved toxins carry-over in waste water can create a sewer
or stream pollution problem if it is not properly addressed.
Material characteristics can influence disposal problems.
Packing and bridging of dry materials in dust hoppers, and
floating or slurry forming characteristics in wet collectors are
examples of problems that can be encountered.
In addition, waste materials originating from air pollution
control devices are hazardous waste as described by U.S. regulators unless they can be proven otherwise.
8.3
DUST COLLECTOR TYPES
The four major types of dust collectors for particulate contaminants are electrostatic precipitators, fabric collectors, wet
collectors, and mechanical collectors.
8.3.1 Electrostatic Precipitators. In electrostatic precipitation, a high potential electric field is established between discharge and collecting electrodes of opposite electrical charge.
The discharge electrode is a small cross-sectional area, such as
a wire or a metal bar, and the collection electrode is large in
surface area such as a plate.
The gas to be cleaned passes through an electrical field that
develops between the electrodes. At a critical voltage, the gas
molecules are separated into positive and negative ions. This
is called ionization and takes place near the surface of the discharge electrode. Ions having the same polarity as the discharge electrode attach themselves to neutral particles in the
gas stream as they flow through the precipitator. These
charged particles are then attracted to a collecting plate of
opposite polarity. Upon contact with the collecting surface,
dust particles lose their charge and then are removed by washing, vibration, or gravity.
The electrostatic process consists of:
1) Ionizing the gas;
2) Charging the dust particles;
3) Transporting the particles to the collecting surface;
8-4
Industrial Ventilation
Air Cleaning Devices
8-5
8-6
Industrial Ventilation
Air Cleaning Devices
8-7
8-8
Industrial Ventilation
4) Neutralizing, or removing the charge from the dust
particles; and
5) Removing the dust from the collecting surface.
The two basic types of electrostatic precipitators are Cottrell
(single-stage) and Penny (two-stage) (Figures 8-4 and 8-5).
The Cottrell single-stage precipitator (Figure 8-4) combines
ionization and collection in a single stage. Because it operates
at ionization voltages from 40,000 to 75,000 volts DC, it may
also be called a high voltage precipitator and is used extensively for heavy duty applications such as utility boilers, larger
industrial boilers, and cement kilns. Some precipitator designs
use sophisticated voltage control systems and rigid electrodes
instead of wires to minimize maintenance problems.
The Penny two-stage precipitator (Figure 8-5) uses DC voltages from 11,000 to 15,000 for ionization and is frequently
referred to as a low voltage precipitator. Its use is limited to low
inlet concentrations, normally not exceeding 0.025 grains per
cubic foot [0.057 g/m3]. It can be the most practical collection
technique for the many condensable hydrocarbon applications
where an initially clear exhaust stack turns into a visible emission as vapor condenses. Some applications include plasticizer
ovens, forge presses, die-casting machines, and various welding operations. Care should be taken to keep the precipitator
inlet temperature low enough to ensure that condensation has
already occurred.
For proper results the inlet gas stream should be evaluated
and treated where necessary to provide proper conditions for
ionization. For high-voltage units a cooling tower is sometimes necessary. Low voltage units may use wet scrubbers,
evaporative coolers, heat exchangers, or other devices to condition the gas stream for best precipitator performance.
The pressure drop of an electrostatic precipitator is extremely low, usually less than 1 "wg [250 Pa]; therefore, the energy
requirement is significantly less than for other techniques.
A modified style of electrostatic collector is used for sticky
submicron aerosol particulate and incorporates some properties
of wet scrubbers and ESPs. It utilizes a continuous coating of
the collection plates with water to cause particulate to collect
on the water surface instead of sticking to the collection plates
themselves. Wet electrostatic precipitation (WESP), once considered experimental, has proven itself a very viable alternative
for some difficult particulate. As with scrubbers, water waste
treatment is a significant issue and wastewater treatability
should be a consideration for every application.
8.3.2 Fabric Collectors. Fabric collectors remove particulate by straining, impingement, interception, diffusion, and
electrostatic charge. The fabric may be constructed of any
fibrous material, either natural or man-made, and may be spun
into a yarn and woven or felted by needling, impacting, or
bonding. Woven fabrics are identified by thread count and
weight of fabric per unit area. Non-woven (felts) are identified
by thickness and weight per unit area. Regardless of construc-
tion, the fabric represents a porous mass through which the gas
is passed unidirectionally so that dust particles are retained on
the dirty side and the cleaned gas passes through.
8.3.2.1 Fabric Filter Efficiencies. Figure 8-15 shows typical filtration efficiency expectations for wet scrubbers, cyclone
collectors and electrostatic precipitators. Note that reversepulse collectors, shakers, and cartridge-style collectors are not
included on the chart. Such media collectors tend to achieve
very high and very similar levels of “seasoned” efficiency
because their efficiency is enhanced by the development of a
“dust cake” on the surface of their filter media.
During steady-state operation a dust cake deposits and
enhances efficiency of the filter media until increased flow
restriction requires disruption and removal of the ‘plugged’
dust cake. Once the dust cake is dislodged there is a brief period of time as a fresh dust cake re-deposits when the media
itself must provide filtration efficiency. The efficiency performance of a dust cake is similar for any collector handling a
similar dust, so the overall efficiency of a collector is impacted
more significantly by other variables such as the frequency at
which cleaning is required, the rate that the dust cake develops,
the filter media used, and/or the mechanical integrity of the
collector including the integrity of the seal between filter
media and the collector.
Properly maintained and conditioned filter media collectors
achieve average efficiencies well in excess of the 99% mass
efficiency across various particle sizes suggested for collection
methods in Figure 8-15. As an example: the total particulate
emissions for the media collector shall average no more than
0.002 grains per dscf (less than 5 mg/m3) over the effective life
of the filters. Improperly maintained media filters, unfortunately, are commonplace and efficiencies may vary directly
with the care the collector receives.
Because media based collectors develop dust cakes on the
surface of the media, the concentrations of dust present at the
media when the dust cake is disrupted will be much higher
than concentrations of dust in the inlet duct to the collector. As
a consequence, filtration efficiency expectations based on dust
concentrations in the inlet duct can be very misleading when
stating actual collector filtration performance. A more accurate
method of establishing performance expectations for media
collectors is to state an acceptable outlet mass emission level
rather than an efficiency for the collector. While collectors and
media selections can achieve very low outlet emissions, it is
important to confirm that there are field test methods available
to verify such performance before establishing low emission
expectations.(8.15)
The ability of the fabric to pass air is defined as permeability
and is measured by the cubic feet of air that pass through one
square foot of fabric each minute at a pressure drop of 0.5 "wg
[125 Pa]. Permeability values for commonly used fabrics
range from 25 to 40 acfm [0.012 to 0.019 am3/s].
A non-woven (felted) fabric is initially more efficient than a
Air Cleaning Devices
8-9
8-10
Industrial Ventilation
woven fabric of identical weight because the void areas or
pores in the non-woven fabric are smaller. A specific type of
fabric can be made more efficient by using smaller fiber diameters, a greater weight of fiber per unit area and by packing the
fibers more tightly. For non-woven construction, the use of
finer needles for felting also improves initial efficiency. While
any fabric is made more efficient by these methods, the cleanability and permeability are reduced. A highly efficient fabric
that cannot be cleaned represents an excessive resistance to airflow and is not an economical engineering solution. Final fabric selection is generally a compromise between efficiency and
permeability.
Over the past 30 years, chemically inert membrane laminates of extended polytetrafluoroethylene (PTFE) or Teflon®
have shown value because of enhanced particulate release and
ultra-high efficiencies. Difficult particulate such as metal fumes
or high temperatures are a good match for PTFE membrane
technologies. However, condensable hydrocarbons and oils
will foul the membranes (Table 8-1).
Choosing a fabric with better cleanability or greater permeability but lower inherent efficiency is not as detrimental as it
may seem. The efficiency of the fabric as a filter is meaningful
only when new fabric is first put into service. Once the fabric
has been in service any length of time, collected particulate in
contact with the fabric acts as a filter aid, defining the real collection efficiency. Compliance testing should never be attempted on new filters until they have been seasoned in service.
Depending on the amount of particulate and the time interval
between fabric reconditioning, it may well be that virtually all
filtration is accomplished by the previously collected particulate
— or dust cake — as opposed to the fabric itself. Even immediately after cleaning, a residual and/or redeposited dust cake
provides additional filtration surface and higher collection efficiency than obtainable with new fabric. While the collection
efficiency of new, clean fabric is easily determined by laboratory test and the information is often published, it is not representative of operating conditions. Therefore it is of little importance
in selecting the proper collector. Please note efficiencies found
in Figure 8-15 for seasoned filter efficiencies.
Fabric collectors are not 100% efficient, but well-designed,
adequately sized, and properly operated fabric collectors can
be expected to operate at efficiencies in excess of 99%, and
often as high as 99.9+% on a mass basis. The inefficiency (or
penetration) that does occur is greatest during or immediately
after reconditioning of the media. Fabric collector inefficiency
is frequently a result of by-pass due to damaged fabric, faulty
seals, or sheet metal leaks rather than penetration of the fabric.
Where extremely high collection efficiency is essential, the
fabric collector should be tested for mechanical leaks. In addition, when highly toxic dusts are involved, a designer should
consider the use of secondary absolute filtration (safety monitoring filters) such as HEPA filters (MERV 17 or the like).
Under some circumstances, even highly toxic particulateladen air streams can be recirculated into the workplace (see
Chapter 11, Section 11.8).
The combination of fabric and collected dust becomes
increasingly efficient as the dust cake accumulates on the fabric surface. At the same time, the resistance to airflow (static
pressure across media – DP) increases. Unless the air moving
device is adjusted to compensate for the increased resistance,
the gas flow rate will be reduced. Figure 8-6 shows how efficiency, resistance to flow and flow rate change with time as
dust accumulates on the fabric. The amount of dust collected
on a single square yard of fabric may exceed five pounds per
hour. In many applications the amount of dust cake accumulated in just a few hours will represent a sufficient increase in
static pressure and cause an unacceptable reduction in system
airflow.
In a well-designed fabric collector system the fabric or filter
mat is cleaned or reconditioned with minimal effect to the system airflow. The cleaning is accomplished by mechanical agitation or air motion that frees the excess accumulation of dust
from the fabric surface and leaves a residual or base cake. This
residual dust cake does not have the same characteristics of
efficiency or resistance to airflow as new fabric. As material
accumulates on the filter, pressure drop across the filter and
collector increases. Below are a few ways the system flow can
be maintained for the variable pressure drop.
1) Monitor collector inlet static pressure:
a) Install a static pressure sensor (pressure transducer)
near or at the inlet of the filtration.
b) If the system temperature varies, a temperature
reading should also be used to compensate for the
pressure reading.
c) The target inlet static pressure will need to be determined when the system is operating at the design
flow (a measurement for average velocity must be
taken simultaneously to show that design flow is
achieved).
d) Use an adjustable fan inlet damper or fan motor
VFD to vary the fan flow based on the inlet static
pressure reading, resulting in a constant volumetric
flow rate.
2) Monitor system airflow rate:
a) Install an airflow meter on the clean side of the collector.
b) The airflow meter should compensate for temperature and pressure changes as appropriate for the system.
c) Use an adjustable fan inlet damper or fan motor
VFD to vary the fan flow based on the airflow reading, resulting in a constant volumetric flow rate.
3) Maintain Hood Static Pressure:
a) Install a static pressure sensor (pressure transducer)
at each system hood capable of sending an appropri-
350 [177]
500 [260]
550 [288]
Vinyon(16)
Clevyl(17)
Glass
Fiberglass(19)
Vinyon
Glass
550 [288]
600 [316]
—
550 [288]
E
E
F
E
E
E
G
E
G
E
F
G
G
G
Dry Heat
E
E
F
E
E
E
F
P
G
E
F
G
F
G
Moist Heat
P
P
F
P–F
P–F
P–F
E
G
E
E
F
G
G
F
Abrasion
P
P
G
G
G
G
E
G
E
E
P–F
G
E
G
Shaking
Resistance to Physical Action
G
F
G
G
G
G
G
E
E
E
G
E
E
G
Flexing
G
E
E
E
E
E
E
P–F
P
P–F
G
G
G
P
G
E
E
E
E
E
E
G
F
E
G
G
G
G
Mineral Acid Organic Acid
G
F
G
E
E
E
E
F
G
G
G
F
F
F
Alkalies
E
E
G
E
E
E
G
G
F
G
G
G
G
F
Oxidizing
Resistance to Chemicals
G
E
P
E
E
E
G
E
E
E
G
E
E
E
Solvents
(1) Du Pont; (2) Celanese; (3) Beaunit; (4) Eastman; (5) American Enka; (6) Chemstrand; (7) American Cyanamid; (8) Farbenfabriken Bayer AG; (9) Dow Chemical; (10) Union Carbide; (11) Allied Chemical; (12) Firestone;
(13) Hercules; (14) Alamo Polymer; (15) National Plastic; (16) FMC; (17) Societe Rhovyl; (18) Lenzing; (19) Huyglas
**Registered Trademarks
*E = excellent; G = good; F = fair; P = poor
Fiberglass
500 [260]
—
450 [232]
Rastex
550 [288]
250 [121]
200 [93]
500 [260]
580 [304]
—
450 [232]
—
500 [260]
225 [107]
400 [204]
Expanded
PFTE
Teflon
Teflon
(Fluorocarbon) TFE(1)
Teflon
FEP(1)
Polyimide
P-84(18)
Polypropylene Herculon(13)
Reevon(14)
Vectra(15)
Nylon
6,6(1,2,6)
Nylon 6(11,5,12)
Nomex(11)
Nylon
(Polyamide)
160 [71]
285 [140]
275 [135]
Dynel(10)
Verel(4)
Modacrylic
—
275 [135]
Orlon(1)
Acrilan(6)
Creslan(7)
Dralon T(8)
Zefran(9)
Acrylic
—
Intermittent
180 [86]
Continuous
Max. Temp. F [C]
Cotton
Dacron(1)
Fortrel(2)
Vycron(3)
Kodel(4)
Enka
Polyester(5)
Example
Trade Name
Fabrics**
Cotton
Polyester
Generic
Names
TABLE 8-1. Characteristics of Filter Fabrics*
Air Cleaning Devices
8-11
8-12
Industrial Ventilation
Air Cleaning Devices
ate signal to the control sensing system.
b) Utilize a programmable logic controller (PLC) that
will allow a minimum setpoint for each hood Static
Pressure (SPh) to be maintained.
c) Test each branch to assure proper flow and utilize
the associated measured hood SPh readings as minimum set points for the PLC to maintain.
d) Use an adjustable fan inlet damper or fan motor
VFD to vary the fan flow to maintain the minimum
hood static pressure at all hoods, resulting in a constant volumetric flow rate.
e) Test periodically (at least quarterly) to insure that
flows and hood static pressures continue to correlate.
Commercially available fabric collectors employ fabric
configured as bags or tubes, envelopes (flat bags), rigid elements, or pleated cartridges. Most of the available fabrics,
whether woven or non-woven, are employed in either bag or
envelope configuration. The pleated cartridge arrangement
typically uses a paper-like fiber in either a cylindrical or panel
configuration. It features extremely high efficiency on light
concentrations. Earlier designs employed cellulose based
media. Today, more conventional media, such as polypropylene or spun-bonded polyester, are frequently used.
The variable design features of the many fabric collectors
available are:
1) Type of fabric (woven or non-woven),
2) Fabric configuration (bags or tubes, envelopes,
cartridges),
3) Intermittent or continuous service,
4) Type of reconditioning (shaker, pulse-jet/reverse-air),
and
5) Housing configuration (single compartment, multiple
compartment).
At least two of these features will be interdependent. For
example, non-woven (felted) fabrics are more difficult to
recondition and, therefore, require high-pressure cleaning.
A fabric collector is selected for its mechanical, chemical,
and thermal characteristics. Table 8-1 lists those characteristics
for some common filter fabrics.
Fabric collectors are sized to provide a sufficient area of filter media to allow operation without excessive pressure drop.
The amount of filter area required depends on many factors,
including:
1)
2)
3)
4)
5)
6)
Release characteristics of dust,
Porosity of dust cake,
Concentration of dust in carrier gas stream,
Type of fabric and surface finish, if any,
Type of reconditioning,
Reconditioning interval,
8-13
7) Airflow pattern within the collector,
8) Temperature and humidity of gas stream,
9) Particle size (aerodynamic), and
10) Collected material characteristics.
Because of the many variables and their range of variation,
fabric collector sizing is a judgment based on experience. The
sizing is usually made by the equipment manufacturer, but at
times may be specified by the user or a third party. Where no
experience exists, a pilot installation is most often used to
determine proper filter type and cloth area.
The sizing or rating of a fabric collector is expressed in
terms of airflow rate versus fabric media area. The resultant
ratio is called air-to-cloth (A/C) ratio with units of acfm per
square foot of fabric. This ratio represents the average velocity
of the gas stream through the filter media. The expression filtration velocity is used synonymously with A/C ratio for rating
fabric collectors. For example, an A/C ratio of 7:1 (7 acfm/sq
ft) is equivalent to a filtration velocity of 7 fpm. In metric values the equivalent would be expressed in m/s. In that case,
0.04 am3/s/m2 is a filtration velocity of 0.04 m/s.
Table 8-2 compares the various characteristics of fabric collectors. The different types will be described in detail later.
The first major classification of fabric collectors is intermittent or continuous duty. Intermittent duty fabric collectors
cannot be reconditioned while in operation. By design, they
require that the gas flow be interrupted while the fabric is agitated to free the accumulated dust cake. Continuous duty collectors do not require shut down for reconditioning.
Shaker Fabric Collectors: Intermittent duty fabric collectors may use a tube, cartridge, or envelope configuration of
woven fabric and will generally employ shaking or vibration
for reconditioning. Figure 8-8 shows both tube and envelope
shaker collector designs. For the tube type, dirty air enters the
open bottom of the tube and dust is collected on the inside of
the fabric. The bottoms of the tubes are attached to a tube sheet
and the tops are connected to a shaker mechanism. Since the
gas flow is from inside to outside, the tubes tend to inflate during operation and no other support of the fabric is required
(Figure 8-7).
Gas flow for envelope type collectors is from outside to
inside, therefore, the envelopes should be supported during
operation to prevent collapsing. This is normally done by
inserting wire mesh or fabricated wire cages into the
envelopes. The opening of the envelope from which the
cleaned air exits is attached to a tube sheet and, depending on
design, the other end may be attached to a support member or
cantilevered without support. The shaker mechanism may be
located in either the dirty air or cleaned air compartments.
The airflow should be stopped periodically (usually at 2- to
6-hour intervals) to recondition the fabric. Figure 8-6 illustrates the system airflow characteristics of an intermittent-duty
fabric collector. As dust accumulates on the fabric, resistance
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Industrial Ventilation
TABLE 8-2. Summary of Fabric Type Collectors and Their Characteristics
to flow increases and airflow decreases until the fan is turned
off and the fabric reconditioned. Variations in airflow due to
changing pressure losses is sometimes a disadvantage and,
when coupled with the requirement to periodically stop the airflow, may preclude the use of intermittent collectors.
Reconditioning seldom requires more than two minutes but
should be done without airflow through the fabric. If reconditioning is attempted with air flowing it will be less effective
and the flexing of the woven fabric will allow a substantial
amount of dust to escape to the ambient atmosphere.
The filtration velocity for large intermittent duty fabric collectors seldom exceeds 6 fpm [0.03 m/s] and normal selections
are in the 2 fpm to 4 fpm [0.01 m/s to 0.02 m/s] range. Lighter
dust concentrations and the ability to recondition more often
allow the use of higher filtration velocities. Ratings are usually
selected so that the pressure drop across the fabric will be in
the 2 to 5 "wg [500 to 1250 Pa] range between start and end of
operating cycle.
With multiple-section, continuous-duty, automatic fabric
collectors, the disadvantage of stopping the airflow to permit
fabric reconditioning and the variations in airflow with dust
cake build-up can be overcome. The use of sections or compartments allows continuous operation of the exhaust system
because automatic dampers periodically remove one section
from service for fabric reconditioning while the remaining
compartments handle the total gas flow. The larger the number
of compartments, the more constant the pressure loss and airflow. Either tubes or envelopes may be used and fabric recon-
ditioning is usually accomplished by shaking or vibrating.
Figure 8-6 shows airflow versus time for a multiple-section
collector. Each individual section or compartment has an airflow versus time characteristic like that of the intermittent collector, but the total variation is reduced because of the multiple
compartments. Note the more constant airflow characteristic of
the five-compartment unit as opposed to the three-compartment design. Since an individual section is out of service only
a few minutes for reconditioning and remaining sections handle the total gas flow during that time, it is possible to clean the
fabric more frequently than with the intermittent type. This permits the multiple-section unit to handle higher dust concentrations. Compartments are reconditioned in fixed sequence with
the ability to adjust the time interval between cleaning of individual compartments.
One variation of this design is the low-pressure, reverse-air collector which does not use shaking for fabric reconditioning.
Instead, a compartment is isolated for cleaning and the tubes collapsed by means of a secondary blower, which draws air from the
compartment in a direction opposite to the primary airflow. This
is a gentle method of fabric reconditioning and was developed primarily for the fragile glass cloth used for high temperature operation, but is now commonplace in the woodworking industry and
other industries where clean, dry, compressed air is not readily
available. The reversal of airflow and tube deflation is accomplished gently to avoid damage to the glass fibers. The control
sequence usually allows the deflation and re-inflation of tubes several times for complete removal of excess dust. Tubes are 6 to 11
Air Cleaning Devices
8-15
8-16
Industrial Ventilation
inches [150 to 280 mm] in diameter and can be as long as 30 feet
[9.1 m]. For long tubes, stainless steel rings may be sewn on the
inside to help break up the dust cake during deflation. A combination of shaking and reverse airflow has also been utilized.
When shaking is used for fabric reconditioning, the filtration velocity usually is in the 1 to 4 fpm [0.005 to 0.02 m/s]
range. Reverse air, collapse type reconditioning generally
necessitates lower filtration velocities since reconditioning is
not as complete. They are seldom rated higher than 3 fpm
[0.015 m/s]. The air to cloth ratio or filtration velocity is based
on net cloth area available when one or more compartments
are out of service for reconditioning.
See also Section 8.7 for Unit Collectors that also employ the
Shaker Fabric collector cleaning method. These collectors are
for small systems or single exhaust source applications.
Reverse Pulse Jet Fabric Collectors: Reverse-jet, continuous-duty, fabric collectors may use envelopes or tubes of nonwoven (felted) fabric, pleated cartridges of non-woven mat
(paper-like) in cylindrical or panel configuration, or rigid elements such as sintered polyethylene. They differ from the low
pressure reverse air type in that they employ a brief burst of
high pressure air to recondition the fabric. Woven fabric is not
used because it allows excessive dust penetration during
reconditioning. The most common designs use compressed air
at 80 to 100 psig [470 to 590 kPa], while others use an integral
blower at a lower pressure but higher secondary flow rate.
Those using compressed air are generally called pulse-jet collectors and those using pressure blowers are called fan-pulse
collectors.
All designs collect dust on the outside and have airflow from
outside to inside the fabric. All recondition the media by introducing the pulse of cleaning air into the opening where cleaned
air exits from the tube, envelope, or cartridge. In many cases, a
venturi shaped fitting is used at this opening to provide additional cleaning by inducing additional airflow. The venturi also
directs or focuses the cleaning pulse for maximum efficiency.
Figure 8-8 shows typical pulse-jet collectors. Under normal
operation (airflow from outside to inside) the fabric shape will
tend to collapse, therefore, a support cage is required. The
injection of a short pulse of high pressure air induces a secondary flow from the clean air compartment in a direction
opposite to the normal airflow. Reconditioning is accomplished by the pulse of high pressure air which stops forward
airflow, then rapidly pressurizes the media, breaking up the
dust cake and freeing accumulated dust from the fabric. The
secondary or induced air acts as a damper, preventing flow in
the normal direction during reconditioning. The entire process,
from injection of the high pressure pulse and initiation of secondary flow until the secondary flow ends, takes place in
approximately one second. Solenoid valves that control the
pulses of compressed air through the diaphragm valves may be
open for a tenth of a second or less. An adequate flow rate of
clean and dry compressed air of sufficient pressure should be
supplied to ensure effective reconditioning.
Reverse-jet collectors normally clean no more than 10% of
the fabric at any one time. Because such a small percentage is
cleaned at any one time and because the induced secondary
flow blocks normal flow during that time, reconditioning can
take place while the collector is in service and without the need
for compartmentalization and dampers. The cleaning intervals
are adjustable and are considerably more frequent than the
intervals for shaker or reverse-air collectors. An individual element may be pulsed and reconditioned as often as once a
minute to every six minutes.
Due to this very short reconditioning cycle, higher filtration
velocities are possible with reverse-jet collectors. However,
accumulated dust that is freed from one fabric surface may
become re-entrained and redeposited on an adjacent surface or
even on the original surface. This phenomenon of redeposition
tends to limit filtration velocity to something less than might
be anticipated with cleaning intervals of just a few minutes.
Laboratory tests(8.1) have shown that, for a given collector
design, redeposition increases with filtration velocity. Other
test work(8.2) indicates clearly that redeposition varies with collector design and especially with flow patterns in the dirty air
compartment. USEPA-sponsored research(8.3) has shown that
superior performance results from downward flow of the dirty
air stream. This downward airflow reduces redeposition since
it aids gravity in moving dust particles toward the hopper.
Interstitial Velocity. Interstitial velocity refers to the upward
air velocity in the collector body between the filter bags. It is
not to be confused with tank velocity or can velocity. The tank
or can velocity is the axial velocity below the filters. Interstitial
velocity as a design parameter is only a consideration in collectors that have an upward flow during the cleaning process. In
addition, newer down-flow collectors reduce this problem significantly.
Depending on the particle size of the collected dust, excessively high interstitial velocities will prevent the dust that is
knocked off the bags by the cleaning system from descending
into the filter hopper for removal. Baghouses have been completely plugged with dust due to high interstitial velocities even
when the air-to-cloth velocity has been properly selected.
One way that vendors design around an interstitial velocity
that is too high is to shorten the bag length. Although tall, skinny baghouses might be desirable from an installation standpoint, interstitial velocities that are too high might dictate going
with the next shorter bag length (and correspondingly larger
baghouse footprint to keep from having filter operating problems). Always consider the interstitial velocity as it compares
to the terminal settling velocity of a particular dust. Submicronsized metal fumes as well as paper fines, textile dust and microscopic feather-like particulate are especially susceptible to this
problem.
Baghouse designers set the target interstitial velocity based
on the aerodynamic particle diameter, particle size and other
Air Cleaning Devices
8-17
8-18
Industrial Ventilation
dust particle properties. Here is the method to calculate a filter’s interstitial velocity so one can determine if a proposed
change (i.e., more airflow or longer bags) is close to the vendor’s recommended velocity:
Given Q = VA then V = Q/A
Vi = Qt/Ai
Vi = Qt/(Af – Ab)
[8.1]
where: Vi = interstitial velocity, fpm [m/s]
Qt = total airflow, acfm [am3/s]
Af = filter housing cross-sectional area, ft2 [m2]
Ab = total bag cross-section area, ft2 [m2]
Ai = interstitial area (Af – Ab), ft2 [m2]
Filtration velocities of 4 to 12 fpm [0.020 H 0.06 m/s] are
normal for reverse-jet collectors. The pleated cartridge type of
reverse-jet collector is limited to filtration velocities in the 7
fpm [0.036 m/s] range and are most often used in the 1 to 3 fpm
[0.005 to 0.015 m/s] range. The pleat configuration may produce very high approach velocities and greater redeposition.
There are many particulate parameters that require a more conservative filtration and interstitial velocity selection. Some are:
•
Hygroscopic (The affinity of a dust to absorb moisture
and become tacky)
•
Abrasive (Cause premature filter or collector failure
and should also be addressed with collector inlet
design)
EXAMPLE PROBLEM 8-1 (From Figure 8-9) (IP Units)
A baghouse has the following size and airflow:
Qt
= 3,000 acfm
Af
= 6′ H 6′ = 36 ft2
Ab
= 49 bags at 0.1963 ft2 or 9.62 ft2
What is interstitial velocity in baghouse?
Vi = Qt/(Ai); 3,000/(36.0 – 9.62) = 3,000/26.4 = 114 feet per
minute
•
Aerodynamic particle diameter (Is the particle more
like a feather or a solid sphere?)
•
Small in size (Typically finer particulate causes more
filter plugging and an inability to recover, especially
particulate smaller than 3 microns in diameter)
•
Fibrous (Fibrous dust can have particularly low bulk
densities and large aerodynamic particle diameters)
A newer type of dust pulse jet dust collector is now widely
used with success and incorporates an enclosing hood built
onto the dust collector itself. The hybrid could be termed a dust
collection booth and is typically used on applications where it
is difficult to apply an exterior hood. One wall of a hopperless
dust collector is open to the booth and the air is brought
through the booth (and across the worker) at 100 to 150 fpm
[0.50 to 0.76 m/s] (similar to a paint spray booth). Fans are
typically incorporated, pulling the media through and recirculating it into the plant air space directly or through HEPA filters. This dust booth concept has been used with success on
welding, sanding, and cutting materials and is coherent with
the concept of enclosing hoods (Figure 8-10). Additionally, it
does not require large energy considerations such as ducts,
hood entry losses, elbows, etc. However, waste handling is significantly more difficult.
8.3.3 Wet Collectors. Wet collectors (scrubbers) are commercially available in many different designs, with pressure
drops from 1.5 "wg [374 Pa] to as much as 100 "wg [24.9 kPa].
There is a corresponding variation in collector performance.
For well-designed equipment, efficiency depends on the energy
utilized in air to water contact and is independent of operating
principle. Efficiency is a function of total energy input per unit
of gas flow whether the energy is supplied to the air or to the
water. This means that well-designed collectors by different
manufacturers will provide similar efficiency if equivalent
power is utilized. Wet collector efficiency is also impacted by
the collector’s ability to effectively remove particulate laden
free water from the air stream (demisting).
Wet collectors have the ability to handle high-temperature
and moisture-laden gases. The collection of dust in a wetted
form minimizes a secondary dust problem in disposal of collected material. Some dusts represent explosion or fire hazards
when dry. Wet collection minimizes the hazard; however, the
use of water may introduce corrosive conditions within the
collector and freeze protection may be necessary if collectors
are located outdoors in cold climates. Pressure losses and collection efficiency vary widely for different designs.
Wet collectors, especially the high-energy types, are frequently the solution to air pollution problems. It should be
realized that disposal of collected material in water without
clarification or treatment may create water pollution problems
and that dried sludges are considered hazardous waste until
otherwise tested.
Wet collectors have one characteristic not found in other
collectors — the inherent ability to humidify. Humidification,
the process of adding water vapor to the air stream through
evaporation, may be either advantageous or disadvantageous
depending on the situation. Where the initial air stream is at an
elevated temperature and not saturated, the process of evaporation reduces the temperature and the volumetric flow rate of
Air Cleaning Devices
8-19
8-20
Industrial Ventilation
FIGURE 8-10. Dust containment booth(8.12)
the gas stream leaving the collector. Assuming the fan is to be
selected for operation on the clean air side of the collector, it
may be smaller and will definitely require less power than if
there had been no cooling through the collector. This is one of
the obvious advantages of humidification; however, there are
other applications where the addition of moisture to the gas
stream is undesirable. For example, the exhaust of humid air
to an air-conditioned space normally places an unacceptable
load on the air conditioning system. High humidity can also
result in corrosion of finished goods. Therefore, humidification effects should be considered before designs are finalized.
All scrubbers are gas conditioners, causing intimate contact
between the particulates in the gas and the scrubbing liquid.
The resulting mixture of clean gases and dust (or fume) laden
water droplets must be channeled through a separation section
for the elimination of entrained droplets. Some collectors, such
as the centrifugal type described in this section, utilize internal
components to collect and remove the water droplets. Spray
towers/chambers and packed bed gas absorbers often use
demisting pads or chevron sections constructed of suitable
materials. Wet dynamic, orifice, and venturi scrubbers may use
a centrifugal separator to remove the droplets. Traditional
cyclonic separators operate at an axial velocity up to 600 fpm
[3 m/s] and high velocity compact separators can operate with
axial velocities up to 1,200 fpm [1.5 m/s], resulting in a much
smaller separator.
Scrubbing water must be discharged from the scrubber to
prevent solids from building up in the scrubbing water of particulate scrubbers. Fresh water is also needed for fume scrubbers (packed tower absorbers) to remove salts formed through
chemical neutralization of the scrubbing water and to ensure
gas absorption efficiency is maintained. The scrubbing water
can be supplied as once-through and treated remotely from the
scrubber. Recycle systems are also common to minimize water
usage. A constant bleed should be maintained on the recycle
system and make-up water will be required to replace the discharged water and any water evaporated in the scrubber.
Chamber or Spray Tower: Chamber or spray tower collectors consist of a round or rectangular chamber into which
water is introduced by spray nozzles. There are many variations of design, but the principal mechanism is impaction of
dust particles on the liquid droplets created by the nozzles.
These droplets are separated from the air stream by centrifugal
force or impingement on water eliminators.
The pressure drop is relatively low (approximately 0.5 to
1.5 "wg [125 to 375 Pa]), but water pressures range from 10 to
400 psig [60 to 2400 kPa]. In general, this type of collector utilizes low pressure supply water and operates in the lower effi-
Air Cleaning Devices
ciency range for wet collectors. Where water is supplied under
high pressure, as with fog towers, collection efficiency can
reach the upper range of wet collector performance.
For conventional equipment, water requirements are reasonable, with a maximum of about 5 gpm per thousand acfm
of saturated gas [approx. 650 liters per thousand am3 saturated
gas]. Fogging types using high water pressure may require as
much as 10 gpm per thousand acfm of saturated gas [approx.
1300 liters per thousand am3 saturated gas]. The fogging
scrubbers are often used in high temperature air streams and
high pressure sprays create a fog whereby particulate and
water vapor/mist interact to grow the water particle into a
droplet of larger size that can be readily removed from the gas
stream.
Packed Towers: Packed towers (Figure 8-11) are essentially
contact beds through which gases and liquid pass concurrently,
counter-currently, or in cross-flow. They are used primarily for
applications involving gas, vapor, and mist removal. These
collectors will capture solid particulate matter but they are not
used for that purpose because dust plugs the packing and
requires unreasonable maintenance. For additional information on packed towers, see Section 8.6.1.
Wet Centrifugal Collectors: Wet centrifugal collectors
(Figure 8-12) comprise a large portion of the commercially
available scrubber designs. This type utilizes centrifugal force
to accelerate the dust particle and impinge it upon a wetted collector surface. Water rates are usually 2 to 5 gpm per thousand
acfm of saturated gas cleaned [approx. 260 to 650 liters per
thousand am3 saturated gas]. Water distribution can be from
nozzles, gravity flow or induced water pickup. Pressure drop
is in the 2 to 6 "wg [500 – 1500 Pa] range.
As a group, these collectors are more efficient than the
chamber type. Some are available with a variable number of
impingement sections. A reduction in the number of sections
results in lower efficiency, lower cost, less pressure drop, and
smaller space. Other designs contain multiple collecting tubes.
For a given airflow rate, a decrease in the tube size provides
higher efficiency because the centrifugal force is greater.
Wet Dynamic Precipitator: Sometimes called a wet fan, the
wet dynamic precipitator (Figure 8-13) is a combination fan
and dust collector. Dust particles in the dirty air stream
impinge upon rotating fan blades wetted with spray nozzles.
The dust particles impinge into water droplets and are trapped
along with the water by a truncated metal cone surrounding the
impeller while the cleaned air makes a turn of 180 degrees and
escapes from the front of the specially shaped impeller blades.
Dirty water from the water cone goes to the water and sludge
outlet and the cleaned air goes to an outlet section containing
a water elimination device. Water rates are usually 0.5 to 1.0
gpm per thousand acfm of saturated gas cleaned [approx. 65 to
130 liters per thousand am3 saturated gas].
Orifice Type Scrubber: In this group of wet collector designs
(Figure 8-13) the airflow through the collector is brought in
8-21
contact with a sheet of water in a restricted passage. Water flow
may be induced by the velocity of the air stream or maintained
by pumps and weirs. Pressure losses vary from 1 "wg [250 Pa]
or less for a water wash paint booth to a range of 2.5 to 11 "wg
[625 to 2750 Pa] for most of the industrial designs. Pressure
drops as high as 20 "wg [5 kPa] are used with some designs
intended to collect very small particles.
Venturi Scrubber: This collector (Figure 8-12) uses a constricted throat to establish throat velocities considerably higher
than those used by the orifice type scrubbers. Gas velocities
through venturi throats may range from 12,000 to 24,000 fpm
[60 to 120 m/s]. Water is supplied by open pipes, injector
tubes, or spray nozzles ahead of the throat at rates from 5 to 15
gpm per thousand acfm [650 to 1950 liters per thousand am3]
of saturated gas.
The primary collection mechanisms of the venturi are
impaction and interception. As is true for all well-designed wet
collectors, collection efficiency increases with higher pressure
drops. Specific pressure drops are obtained by designing for
selected velocities in the throat of the venturi. Some are made
with adjustable throats allowing operation over a range of
pressure drops for a given flow rate or over a range of flow
rates with a constant pressure drop. Systems are available with
pressure drops as low as 5 "wg [1250 Pa] for moderate collection efficiency and as high as 100 "wg [25 kPa] for collection
of extremely fine particles, such as fumes or other aerosols. If
a variable flow or range of pressure drops is to be used, it is
important that the system fan be able to accommodate the
varying conditions.
8.3.4 Dry Centrifugal Collectors. Dry centrifugal collectors separate entrained particulate from an air stream by the
use or combination of centrifugal, inertial, and gravitational
force. Collection efficiency is influenced by:
1) Particle size, weight, and shape. Performance is
improved as particle size and weight become larger and
as the shape becomes more spherical.
2) Collector size and design. The collection of fine dust
with a mechanical device requires equipment designed
to optimize mechanical forces and provide enough residence time for collection of the fine dust.
3) Velocity. Pressure drop through a cyclone collector
increases approximately as the square of the inlet
velocity. Increasing the inlet velocity (or reducing the
cyclone diameter) increases the centrifugal force used
to collect the particle. There is, however, a maximum
velocity that is a function of collector design, where
material carryover and a reduced efficiency will occur.
4) Dust Concentration. Generally, the performance of a
mechanical collector increases as the concentration of
dust becomes greater.
Gravity Separators: Gravity separators (often referred to as
8-22
Industrial Ventilation
Air Cleaning Devices
8-23
8-24
Industrial Ventilation
Air Cleaning Devices
a drop-out box or drop-out chamber) consist of a chamber or
housing in which the velocity of the gas stream is made to drop
rapidly so that dust particles settle out by gravity. Extreme
space requirements and the usual presence of eddy currents
nullify this method for removal of anything but extremely
coarse particles.
Inertial Separators: Inertial separators depend on the inability of dust to make a sharp turn because its inertia is much
higher than that of the carrier gas stream. Blades, baffles or
louvers in a variety of shapes are used to require abrupt turns
of 120 degrees or more. Well-designed inertial separators can
remove particles in the 10 to 20 micron range with about 90%
efficiency.
Cyclone Collector: The cyclone collector (Figure 8-14) is
commonly used for the removal of coarse dust, as a precleaner to more efficient dust collectors and/or as a product separator in air conveying systems. Principal advantages are low
cost, low maintenance, and relatively low pressure drops (in
the 3.0 to 6.0 "wg [approx. 760 to 1480 Pa] range). It is not
suitable for the collection of fine particles (Figure 8-15 IP &
SI).
High Efficiency Cyclone Collectors: High efficiency
cyclones (Figure 8-14) are often larger than a traditional
cyclone to provide additional time for fine particulate to be
collected, have a higher pressure drop to exert higher centrifugal forces on the dust particles, or a combination of both.
Improved dust separation efficiency in high efficiency
cyclones has been obtained by:
1) Increasing the inlet velocity. This increases the centrifugal force and the pressure drop.
2) Increasing the cyclone size while maintaining inlet
velocity. This allows for more time for fine particulate
to be collected.
3) Using a number of small diameter cyclones in parallel.
This reduces the body diameter, increasing the centrifugal force. Cyclones handling a smaller gas volume are
more efficient than the same family of cyclones handling a larger gas volume.
4) Placing units in series. Pressure drops of cyclones
installed in series are additive, resulting in higher energy usage. Often the second cyclone in series is a higher
efficiency model than the first cyclone in series. This is
a common arrangement in critical processes where the
second cyclone serves as a redundant cyclone in case
the first cyclone plugs or malfunctions.
While high efficiency centrifugal collectors are not as efficient on small particles as electrostatic, fabric, and wet collectors, their effective collection range is appreciably extended
beyond that of other mechanical devices. Pressure losses of collectors in this group range from 3 to 12 "wg [750 to 3000 Pa].
8.4
8-25
ADDITIONAL AIDS IN DUST COLLECTOR
SELECTION
The collection efficiencies of the five basic groups of air
cleaning devices have been plotted against mass mean particle
size (Figure 8-16). The graphs were found through laboratory
and field testing and were not compiled mathematically. The
number of lines for each group indicates the range that can be
expected for the different collectors operating under the same
principle. Variables (type of dust, aerodynamic particle diameter, velocity of air, water rate, etc.) will also influence the
range for a particular application.(8.15,8.16,8.17)
8.5
CONTROL OF MIST, GAS AND VAPOR
CONTAMINANTS
Previous discussion has centered on the collection of dust
and fume or particulate existing in the solid state. Only the
packed tower was singled out as being used primarily to collect mist, gas, or vapor. The character of a mist aerosol is very
similar, aerodynamically, to that of a dust or fume aerosol, and
the mist can be removed from an air stream by applying the
principles that are used to remove solid particulate.
Standard wet collectors are used to collect many types of
mists. Specially designed electrostatic precipitators are frequently employed to collect sulfuric acid or oil mist. Even fabric and centrifugal collectors, although not the types previously mentioned, are widely used to collect oil mist generated by
high speed machining.
Oil aerosols less than 1 micron in diameter, typically associated with blue haze require special collectors. High-density
fiber-bed filters, electrostatic precipitators (wet and dry) are typically used for this most difficult type of aerosol contaminant.
8.6
GASEOUS CONTAMINANT COLLECTORS
Industrial processes produce tremendous quantities of
gaseous contaminants. The terms gas and vapor are commonly
incorrectly used interchangeably. Matter that takes both the
shape and volume of its container is said to be in a gaseous
state. Gas molecules contain enough energy to continue to
move apart until they bounce off the sides of the container(s)
holding them. The term gas describes those substances that
exist in a gaseous state at room temperature.
Vapor describes a substance that, although in the gaseous
state, is generally a liquid or solid at room temperature. Steam,
the gaseous form of water, is a vapor. Moist air contains water
vapor. Partial pressure relationships described by Dalton’s
Law explain how water vapor and dry air coexist at room temperature and atmospheric pressure. (Refer to Chapter 3,
Section 3.9 for further discussion of psychrometric principles.)
Numerous techniques have been developed to control
gaseous contaminants. The more commonly used techniques
include Absorption, Adsorption, Incineration/Oxidation, and
Biofiltration. Newer control methods include corona reactors,
direct electric arcing, plasma treatment, condensation, and
8-26
Industrial Ventilation
Air Cleaning Devices
8-27
8-28
Industrial Ventilation
TABLE 8-3. Dust Collector Selection Guide
Collector Types Used in Industry
Operation
CERAMICS
a. Raw product handling
b. Fettling
c. Refractory sizing
d. Glaze & vitr. enamel spray
CHEMICALS
a. Material handling
b. Crushing, grinding
c. Pneumatic conveying
d. Roasters, kilns, coolers
Concentration
Note 1
Particle
Sizes
Note 2
Dry Centrifugal
Collector
Wet
Collector
Fabric
Collector
Low-Volt
Electrostatic
Hi-Volt
Electrostatic
light
light
heavy
moderate
fine
fine-medium
coarse
medium
S
S
N
N
O
S
S
O
O
O
O
O
N
N
N
N
N
N
N
N
lightmoderate
moderateheavy
very
heavy
heavy
finemedium
finecoarse
finecoarse
midcoarse
S
O
O
N
N
49
4
O
O
O
N
N
5
O
S
O
N
N
6
O
O
O
N
N
7
medium
fine
mediumcoarse
fine
O
S
S
O
O
O
O
O
O
N
N
N
N
N
N
49
8
9
10
S
O
O
N
N
11
light
moderate
fine
finecoarse
S
S
S
S
O
O
N
N
O
O
12
moderate
varies
fine
coarse
S
S
S
S
O
O
N
N
O
S
13
14
lightmoderate
moderate
fine
N
O
O
N
N
15
S
O
O
N
N
16
N
O
O
N
N
17
N
O
O
N
N
18
O
S
O
O
O
O
O
O
O
O
O
O
N
N
N
N
N
N
N
N
N
N
O
O
S
S
N
N
S
S
N
N
N
N
S
O
S
S
O
O
O
O
N
N
N
N
S
N
N
25
26
27
28
N
O
N
N
N
O
O
O
O
N
N
N
O
O
S
O
O
O
N
N
N
N
N
N
N
O
S
S
N
N
29
30
31
32
33
34
COAL, MINING AND POWER PLANT
a. Material handling
moderate
b. Bunker ventilation
light
c. Dedusting, air cleaning
heavy
d. Drying
FLY ASH
a. Coal burning—chain grate
b. Coal burning—stoker fired
c. Coal burning—pulverized
fuel
d. Wood burning
FOUNDRY
a. Shakeout
b. Sand handling
moderate
finemedium
c. Tumbling mills
heavy
mediumcoarse
d. Abrasive cleaning
moderate- fineheavy
medium
GRAIN ELEVATOR, FLOUR AND FEED MILLS
a. Grain handling
light
medium
b. Grain dryers
light
coarse
c. Flour dust
moderate
medium
d. Feed mill
moderate
medium
METAL MELTING
a. Steel blast furnace
heavy
varied
b. Steel open hearth
moderate
finecoarse
c. Steel electric furnace
light
fine
d. Ferrous cupola
moderate
varied
e. Non-ferrous reverberatory varied
fine
f. Non-ferrous crucible
light
fine
METAL MINING AND ROCK PRODUCTS
a. Material handling
moderate
fine-medium
b. Dryers, kilns
moderate
medium-coarse
c. Rock dryer
moderate
fine-medium
d. Cement kiln
heavy
fine-medium
e. Cement grinding
moderate
fine
f. Cement clinker cooler
moderate
coarse
See
Remark No.
1
2
3
49
19
20
21
22
49
23
24
Air Cleaning Devices
8-29
TABLE 8-3 (Cont.). Dust Collector Selection Guide
Collector Types Used in Industry
Operation
Concentration
Note 1
Particle
Sizes
Note 2
METAL WORKING
a. Production grinding,
light
coarse
scratch brushing, abrasive
cut off
b. Portable and swing frame
light
medium
c. Buffing
light
varied
d. Tool room
light
fine
e. Cast iron machining
moderate
varied
PHARMACEUTICAL AND FOOD PRODUCTS
a. Mixers, grinders, weighing, light
medium
blending, bagging,
packaging
b. Coating pans
varied
finemedium
PLASTICS
a. Raw material processing
(See comments under
Chemicals)
b. Plastic finishing
lightvaried
moderate
c. Extrusion
light
fine
RUBBER PRODUCTS
a. Mixers
moderate
fine
b. Batchout rolls
light
fine
c. Talc dusting and dedusting moderate
medium
d. Grinding/buffing
moderate
coarse
WOODWORKING
a. Woodworking machines
moderate
varied
b. Sanding
moderate
fine
c. Waste conveying, hogs
heavy
varied
Dry Centrifugal
Collector
Wet
Collector
Fabric
Collector
Low-Volt
Electrostatic
Hi-Volt
Electrostatic
See
Remark No.
O
O
O
N
N
49
35
S
O
S
O
O
O
S
O
O
O
S
O
N
N
N
S
N
N
N
N
36
37
38
O
O
O
N
N
39
N
O
O
N
N
40
O
S
O
N
N
49
41
S
S
O
N
N
42
N
S
N
O
N
S
S
S
O
O
O
S
O
S
S
O
O
N
S
N
N
N
N
N
N
O
S
O
S
S
S
O
O
O
N
N
N
N
N
N
49
43
44
45
49
46
47
48
Note 1: Light: less than 2 gr/ft3 [4.6 g/m3]; Moderate: 2 to 5 gr/ft3 [4.6 – 14.4 g/m3]; Heavy: 5 gr/ft3 [14.4 g/m3] and up.
Note 2: Fine: 50% less than 5 microns; Medium: 50% 5 to 15 microns; Coarse: 50% 15 microns and larger.
Note 3: O = often; S = seldom; N = never.
Remarks Referred to in Table 8-3
1. Dust released from bin filling, conveying, weighing, mixing,
10. Heavy loading suggests final high efficiency collector for all
except very remote locations.
pressing, forming. Refractory products, dry pan and screen
operations more severe.
11. Difficult problem but collectors will be used more frequently
with air pollution emphasis.
2. Operations found in vitreous enameling, wall and floor tile,
12. Public nuisance from boiler blow-down indicates collectors are
pottery.
needed.
3. Grinding wheel or abrasive cut-off operation. Dust abrasive.
13. Large installations in residential areas require electrostatic in
4. Operations include conveying, elevating, mixing, screening,
addition to dry centrifugal.
weighing, packaging. Category covers so many different
14.
Cyclones
used as spark arresters in front of fabric collectors.
materials that recommendation will vary widely.
15. Hot gases and steam usually involved.
5. Cyclone and high efficiency centrifugals often act as primary
16. Steam from hot sand, adhesive clay bond involved.
collectors followed by fabric or wet type.
17. Concentration very heavy at start of cycle.
6. Cyclones used as product collector followed by fabric arrester
18. Heaviest load from airless blasting due to higher cleaning
for high overall collection efficiency.
speed. Abrasive shattering greater with sand than with grit or
7. Dust concentration determines need for dry centrifugal; plant
shot. Amounts removed greater with sand castings, less with
location, product value determines need for final collectors.
forging scale removal, least when welding scale is removed.
High temperatures are usual and corrosive gases not unusual.
19. Operations such as car unloading, conveying, weighing,
8. Conveying, screening, crushing, unloading.
storing.
9. Remove from other dust producing points. Separate collector
20. Collection equipment expensive but public nuisance complaints
usually.
becoming more frequent.
8-30
Industrial Ventilation
Remarks Referred to in Table 8-3 (continued)
21. Operations include conveyors, cleaning rolls, sifters, purifiers,
bins and packaging.
36. Linty particles and sticky buffing compounds can cause
pluggage and fire hazard in dry collectors.
22. Operations include conveyors, bins, hammer mills, mixers,
feeders and baggers.
37. Unit collectors extensively used, especially for isolated
machine tools.
23. Primary dry trap and wet scrubbing usual. Electrostatic is
added where maximum cleaning required.
38. Dust ranges from chips to fine floats including graphitic carbon.
Low voltage ESP applicable only when a coolant is used.
24. Use of this technique declining.
39. Materials vary widely. Collector selection depends on salvage
value, toxicity, sanitation yardsticks.
25. Air pollution standards will probably require increased usage
of fabric arresters.
26. Continuous coating (dry-scrubbing) of emissions is
recommended for scrap remelting.
27. Zinc oxide loading heavy during zinc additions. Stack
temperatures high.
28. Zinc oxide plume can be troublesome in certain plant locations.
29. Crushing, screening, conveying involved. Wet ores often
introduce water vapor in exhaust air.
30. Dry centrifugals used as primary collectors, followed by final
cleaner.
31. Industry is aggressively seeking commercial uses for fines.
32. Collectors usually permit salvage of material and also reduce
nuisance from settled dust in plant area.
33. Salvage value of collected material high. Same equipment
used on raw grinding before calcining.
34. Coarse abrasive particles readily removed in primary collector
types.
35. Roof discoloration, deposition on autos can occur with cyclones
and less frequently with high efficiency dry centrifugal. Heavy
duty air filters sometimes used as final cleaners.
photochemical oxidation.
8.6.1 Absorption. Absorption is a mass transfer process
where transfer occurs through a phase boundary and the collected molecule is held within the absorbing medium.
Absorbers remove soluble or chemically reactive gases from
the gas stream through intimate contact with a suitable liquid
so that one or more of the gas stream components will dissolve
in the liquid. While all designs utilize intimate contact between
the gaseous contaminant and the absorbent, they vary widely
in configuration and performance. Removal may be by
absorption if the gas solubility and vapor pressure promote
absorption or chemical reaction. There are both dry and wet
absorbers. In wet absorbers, water is the most frequently used
absorbent, but additives are frequently required and occasionally other chemical solutions should be used. Typical wet
absorber designs include packed scrubbers, staged devices,
and high energy contactors (venturi scrubbers).
Packed Scrubbers: Variants of the packed scrubber are
available in four configurations. They are the horizontal cocurrent scrubber, the vertical cocurrent scrubber, the crossflow
scrubber, and the countercurrent scrubber (Figure 8-11). The
40. Controlled temperature and humidity of supply air to coating
pans makes recirculation desirable.
41. Plastic manufacture allied to chemical industry and varies with
operations involved.
42. Operations and collector selection similar to woodworking.
See Item 13.
43. Concentration is heavy during feed operation. Carbon black
and other fine additions make collection and dust-free
disposal difficult.
44. Salvage of collected material often dictates type of high
efficiency collector.
45. Fire hazard from some operations should be considered.
46. Bulking material. Collected material storage and bridging from
splinters and chips can be a problem.
47. Dry centrifugals not effective on heavy concentration of fine
particles from production sanding.
48. Dry centrifugal collectors required. Wet or fabric collectors
may be used for final collectors.
49. See NFPA publications for fire and explosion hazards, e.g.,
zirconium, magnesium, aluminum, woodworking, plastics, etc.
horizontal cocurrent scrubber depends on the gas velocity to
carry the liquid into the packed bed and operates as a wetted
entrainment separator with limited gas and liquid contact time.
A vertical cocurrent scrubber may be operated at pressure
drops of 1 to 3 inches of water [250 to 750 Pa] per foot [0.3 m]
of packing depth. Contact time is a function of packing depth
in this configuration.(8.4)
Crossflow scrubbers use a horizontal gas stream movement
with the liquid scrubbing medium flowing down through the
gas stream. Absorption efficiency for this design is generally
somewhere between that of cocurrent and countercurrent flow
scrubbers.
Countercurrent scrubbers have the gas flowing up through a
downward liquid flow. The efficiency of countercurrent scrubbers is maximized because the exit gas is in contact with the
fresh scrubbing liquor where the highest driving forces exist to
aid the mass transfer process. Packed towers are countercurrent scrubbers. The packed tower scrubber (previously discussed in Section 8.3.3) consists of a cylindrical shell, a
packed section held on a support plate, a liquid distributor,
possibly a liquid redistributor, access manholes, gas inlet and
Air Cleaning Devices
8-31
FIGURE 8-16. Characteristics of particles and particle dispersoids (Reprinted with permission from SRI International)
outlet, and possibly a sump with recirculation pump and overflow. There are a wide variety of packing materials available.
Packings providing more surface area per unit volume are generally regarded as superior. There are tradeoffs to consider
when selecting a packing material which will impact the overall equipment height and pressure drop requirements to meet
specific contaminant collection removal characteristics
(Figure 8-11).(8.5)
brick linings, allowing gas temperatures as high as 1,600 F
[871 C] to be handled directly from furnace flues.
Water rates of 5 to 10 gpm per thousand acfm [approx. 650
to 1300 thousand am3] of saturated gas are typical for packed
towers. Water is distributed over V-notched ceramic or plastic
weirs. High temperature deterioration is avoided by using
Staged Scrubbers: Staged or stagewise equipment utilizes a
group of horizontal metal plates arranged in a vertical series
and generally placed in a cylindrical housing. Each horizontal
plate is a stage. The plates can be sieves, bubble type or bal-
The airflow pressure loss for a 4 foot [1.1 m] bed of packing, such as ceramic saddles, will range from 1.5 to 3.5 "wg
[375 to 875 Pa]. The face velocity (velocity at which the gas
enters the bed) will typically be 200 to 600 fpm [approx. 1.0 to
3.0 m/s].
8-32
Industrial Ventilation
lasts. Gas flow is countercurrent to the liquid flow in all cases.
In each of these designs, the liquid is kept on the tray surface
by a dam at the entrance to a downcomer or sealed conduit
allowing overflow liquid to pass to the tray below.(8.6)
These scrubbers are often used to condense or sub-cool the
gas stream. In order to achieve the desired outlet gas temperature, heat must be removed from the recycle water system.
High Energy Scrubbers: High Energy Contactors (Venturi
Scrubbers, Figures 8-12 and 8-13) were also described in
Section 8.3.3. Although used predominately as particulate control devices they can simultaneously function as absorbers.
Venturi scrubbers are cocurrent devices and their absorption
characteristics are maximized when operating at low velocities
with high liquid to gas ratios.
Dry Absorption: Dry absorption systems include dry scrubbers, spray dryers and fluid bed reactors. Dry scrubbers
involve injection of a dry sorbent directly into a process gas
stream. Spray dryers inject a wet sorbent into a hot gas stream
where the liquid evaporates leaving a dry solvent in contact
with the gas. Fluid bed reactors employ a bed of granulated
solvent fluidized within a vessel and the process gas flows
through the fluidized bed. All dry absorption systems should
include an appropriate particulate removal device in order to
remove reaction products, excess sorbent material and particulate matter from the gas stream.
8.6.2 Adsorption. Adsorption is also a mass transfer pro-
cess that removes contaminants by adhesion of molecules of
one phase to the surface or interfaces of a solid second phase.
Relatively weak adsorption, where the forces involved are
intermolecular, is known as van der Waals adsorption. Strong
adsorption, where the forces involved are valence forces, is
known as activated adsorption or chemisorption. No chemical
reaction is involved as adsorption is a physical process that is
normally thought of as reversible. Activated carbon, activated
alumina, silica gel, Fuller’s earth, and molecular sieves are
popular adsorbents.
8.6.3 Incineration/Oxidation. These two terms, incineration
and oxidation, are used interchangeably to describe the process
of combustion. Combustion is a chemical process in which
oxygen reacts with various elements or chemical compounds
resulting in the release of light and heat. The combustion process readily converts volatile organic compounds (VOCs),
organic aerosols, and most odorous materials to carbon dioxide
and water vapor. It is a very effective means of eliminating
VOCs. Typical applications for incineration devices include
odor control, reduction in plume opacity caused by condensable particulate, reduction in reactive hydrocarbon emissions,
and reduction of explosion hazards. The equipment used for
control of gaseous contaminants by combustion may be divided into three categories: thermal oxidizers, direct combustors,
or catalytic oxidizers.
Thermal Oxidizers, or afterburners, may be used where the
contaminant is combustible. The contaminated air stream is
introduced to an open flame or heating device followed by a
residence chamber where combustibles are oxidized producing carbon dioxide and water vapor. Most combustible contaminants can be oxidized at temperatures between 1,000 F
[538 C] and 1,500 F [816 C]. The residence chamber should
provide sufficient dwell time and turbulence to allow complete
oxidation. Thermal oxidizers are often equipped with heat
exchangers where combustion gas is used to preheat the
incoming contaminated gas. If gasoline is the contaminant,
heat exchanger efficiencies are limited to 25 to 35% and preheat temperatures are maintained below 530 F [277 C] to minimize the possibility of ignition occurring in the heat exchanger. Flame arrestors are always installed between the vapor
source and the thermal oxidizer. Burner capacities in the combustion chamber range from 0.5 to 2 M BTU [527.9 to 2111.6
kJ] per hour. Operating temperatures range from 1,400 to
1,600 F [760 C to 871 C], and gas residence times are typically
1 second or less. This condition causes the molecular structure
to break down into carbon dioxide and water vapor.
Regenerative thermal oxidation (RTO) units are distinguished from other thermal incinerators by their ability to
recover heat at high efficiency. RTOs employ two, three, five,
seven, or more chambers that store and recycle heat energy.
RTO technology uses high temperature to convert VOCs into
carbon dioxide and water vapor.
In the RTO, contaminated process air enters a combustion
chamber after being preheated through a ceramic bed, where
the air is raised to a required temperature and held there for a
specified period of time. The heat recovery chambers are outfitted with stoneware or ceramic beds that absorb most of the
heat energy from the combustion chamber. The flow is then
reversed, allowing the next contaminated batch of air to enter
the combustion chamber through the stoneware bed that was
heated from the last batch. The level of heat recovery varies,
depending on the specific design of the system.
Using a flameless thermal oxidation process, VOC-laden
exhaust gas typically enters a single or multiple module RTO.
The VOC gas stream is alternatively directed using valves to
the top or bottom air plenum and is transported through a
porous gravel heat exchange bed. In the gravel media, it is
flamelessly oxidized and converted to carbon dioxide and
water vapor. Reversal of the gas stream keeps the high temperature band centered in the gravel media. For start-up, natural
gas/propane is injected into the heat transfer media to bring the
temperature up to approximately 1,800 F [982 C]. For low
concentration streams of VOC exhaust, supplemental fuel is
needed to maintain the proper oxidation temperature. For
VOC streams above a concentration of 3.8%, the reaction is
self-sustaining. The process attains greater than 98% VOC
destruction and 95% heat recovery.
Direct combustors (flares) differ from thermal oxidizers by
introducing the contaminated gases and auxiliary air directly
into the burner as fuel. Auxiliary fuel, usually natural gas or
oil, is generally required for ignition. It may or may not be
Air Cleaning Devices
required to sustain burning and all of the waste gases react at
the burner.
Catalytic Oxidation: Catalytic oxidation is a more recent
alternative for the treatment of VOCs in air streams. It is very
similar to thermal oxidation, except that with a catalyst present, the same reaction occurs at a lower temperature.
Catalysts are substances that alter the rate of a chemical reaction without themselves being consumed in the reaction.
VOCs are thermally destroyed at temperatures typically ranging from 600 to 1,000 F [316 C to 538 C] by using a solid catalyst. First, the contaminated air is directly preheated (electrically or, more frequently, using natural gas or propane) to
reach a temperature necessary to initiate the catalytic oxidation
of the VOCs. Then the preheated VOC-laden air is passed
through a bed of solid catalysts where the VOCs are rapidly
oxidized.
In most cases, the process can be enhanced to reduce auxiliary fuel costs by using an air-to-air heat exchanger to transfer
heat from the exhaust gases to the incoming contaminated air.
Typically, 50–70% of the heat of the exhaust gases is recovered. Depending on VOC concentrations, the recovered heat
may be sufficient to sustain oxidation without additional fuel.
Catalyst systems used to oxidize VOCs typically use metal
oxides such as nickel oxide, copper oxide, manganese dioxide,
or chromium oxide. Noble metals such as platinum and palladium may also be used.
To use either thermal or catalytic oxidation, the combustible
contaminant concentration should be below the lower explosive limit. Equipment specifically designed for control of
gaseous or vapor contaminants should be applied with caution
when the air stream also contains solid particles. Solid particulate can plug absorbers, adsorbers, and catalysts and, if noncombustible, will not be converted in thermal oxidizers and
direct combustors. In addition, chemicals such as sulfur, silicone, metals or halogens can poison a catalyst.
8.6.4 Biofiltration.(8.7,8.8) The biofiltration process involves
drawing contaminated air through a pretreatment unit to adjust
its temperature and moisture content, and then through a filter
in which the contaminants are transferred to microorganisms
selected for their efficiency in treating those specific contaminants. It is a more recent air pollution control technology suited for cleaning VOCs and other gases such as ammonia and
hydrogen sulfide. These gases are considered responsible for
odors associated with some organic products. Successful and
common applications of biofilters in agricultural facilities,
rendering plants, wastewater treatment plants, chemical,
wood panel manufacturing, and food processing plants.
8.6.5 Other Gaseous Contaminant Controls. The most
commonly used of the lesser known gaseous contaminant control methods referred to above is condensation. It has been
widely used for recovery of and/or removal of gaseous specific
constituents in a bulk gas flow. Specific examples would
include the selective distillation of various hydrocarbons in
8-33
refining processes and the drying of air. In order to remove a
selected contaminant from a gas stream by this method the dew
point of the pollutant should be significantly higher than that of
the non-contaminant gases. This technique has been successfully applied as a control method for removal of some VOCs.
Application of the corona reactor, photochemical oxidation,
direct electric arcing, and plasma treatment techniques are still
somewhat experimental at this date. All of these techniques target VOCs and some inorganic gases such as hydrogen sulfide,
mercaptans, trichloroethylene, and carbon tetrachloride. Air
streams containing both solid particles and gaseous contaminants may require appropriate control devices in series.
8.7
UNIT COLLECTORS
Unit collector is a term usually applied to small fabric collectors having capacities in the 200 to 5,000 acfm [0.1 to 2.5
am3/s] range. They have integral air movers, feature small
space requirements and are simple to install. In most applications, cleaned air is recirculated (exhausted into the
workspace), although discharge ducts may be used if the
added resistance is within the capability of the air mover. One
of the primary advantages of unit collectors is a reduction in
the amount of duct required, as opposed to central systems.
Combustible dust regulations such as NFPA 654 and 664
allow for recirculation of air based on the unit collector meeting a definition of an “enclosureless collector” with further
stipulations.(8.11)
When cleaned air is to be recirculated, a number of precautions are required (see Chapter 11).
Unit collectors are used extensively to fill the need for dust
collection from isolated, portable, intermittently used, or frequently relocated dust producing operations. Typically, a single collector serves one dust source with the energy saving
advantage that the collector would operate only when that particular dust producing machine is in operation.
Figure 8-17 shows a typical unit collector. Usually they are
the intermittent duty, shaker-type in envelope configuration.
Woven fabric is nearly always used. Automatic fabric cleaning
is preferred as manual methods without careful scheduling and
supervision are unreliable.
8.8
DUST COLLECTING EQUIPMENT COST
The variations in equipment cost, especially on an
installed basis, are difficult to estimate. Comparisons can be
misleading if these factors are not carefully evaluated.
8.8.1 Price Versus Capacity. All dust collector prices per
acfm of gas will vary with the gas flow rate. The smaller the
flow rate, the higher the cost per acfm. The break point, where
price per acfm cleaned tends to level off, will vary with the
design. See the typical curves shown in Figure 8-18.
8.8.2 Accessories Included. Careful analysis of components of equipment included is very important. Some collector
designs include exhaust fan, motor, drive, and starter. In other
8-34
Industrial Ventilation
Air Cleaning Devices
8-35
8-36
Industrial Ventilation
Air Cleaning Devices
8-37
designs, these items and their supporting structure should be
obtained by the purchaser from other sources. Likewise, while
dust storage hoppers are integral parts of some dust collector
designs, they are not provided in other types. Duct connections
between elements may be included or omitted. Recirculating
water pumps and/or settling tanks may be required but not
included in the equipment price.
8.9.2 Impingement. When air flows through a filter, it
changes direction as it passes around each fiber. Larger dust
particles, however, cannot follow the abrupt changes in direction because of their inertia. As a result, they do not follow the
air stream and collide with a fiber. Filters using this method are
often coated with an adhesive to help fibers retain the dust particles that impinge on them.
8.8.3 Installation Cost. The cost of installation can equal
or exceed the cost of the collector. Actual cost will depend on
the method of shipment (completely assembled, sub-assembled, or completely knocked down), the location (that may
require expensive rigging), and the need for expensive supporting steel and access platforms. Factory installed media
will reduce installation cost. The cost can also be measurably
influenced by the need for water and drain connections, special or extensive electrical work, and expensive material handling equipment for collection material disposal. Items in the
latter group will often also be variable, decreasing in cost per
acfm as the flow rate of gas to be cleaned increases.
8.9.3 Interception. Interception is a special case of
impingement where a particle is small enough to move with
the air stream but, because its size is very small in relation to
the fiber, makes contact with a fiber while following the tortuous airflow path of the filter. The contact is not dependent on
inertia and the particle is retained on the fiber because of the
inherent adhesive forces that exist between the particle and
fiber. These forces (called van der Waals forces) enable a fiber
to trap a particle without the use of inertia.
8.8.4 Special Construction. Prices shown in any tabulation should necessarily assume standard or basic construction.
The increase in cost for corrosion resisting material, special
high temperature fabrics, insulation, and/or weather protection for outdoor installations can introduce a multiplier of one
to four times the standard cost.
8.9.4 Diffusion. Diffusion takes place on particles so small
that their direction and velocity are influenced by molecular
collisions. These particles do not follow the air stream, but
behave more like gases than particulate. They move across the
direction of airflow in a random fashion. When a particle does
strike a fiber, it is retained by the van der Waals forces existing
between the particle and the fiber. Diffusion is the primary
mechanism used by most extremely efficient filters.
A general idea of relative dust collector cost is provided in
Figure 8-18. The additional notes and explanations included in
these data should be carefully examined before they are used
for estimating the cost of specific installations. For more accurate data, the equipment manufacturer or installer should be
asked to provide estimates or a past history record for similar
control problems utilized. Table 8-4 lists other characteristics
that should be evaluated along with equipment cost.
8.9.5 Electrostatic. A charged dust particle will be attracted
to a surface of opposite electrical polarity. Most dust particles
are not electrically neutral, therefore, electrostatic attraction
between dust particle and filter fiber aids the collection efficiency of all barrier type air filters. Electrostatic filters establish an
ionization field to charge dust particles so that they can be collected on a surface that is grounded or of opposite polarity. This
concept was previously discussed in Section 8.3.1.
Price estimates included in Figure 8-18 are for equipment of
standard construction in normal arrangement. Estimates for
exhausters and dust storage hoppers have been included, as
indicated in Notes 1 and 2, where they are normally furnished
by others.
8.9.6 Disposable Filter Rating. Table 8-5 shows performance versus filter fiber size for several filters. Note that efficiency increases as fiber diameter decreases because more
small fibers are used per unit volume. Note also that low
velocities are used for high efficiency filtration by diffusion.
8.9
SELECTION OF DISPOSABLE-TYPE AIR
FILTRATION EQUIPMENT
Air filtration equipment is available in a wide variety of
designs and capabilities. Performance ranges from a simple
throwaway filter for the home furnace to the clean room in the
electronics industry, where the air should be a thousand times
as clean as in a hospital surgical suite. Selection is based on
efficiency, dust holding capacity, and pressure drop. There are
five basic methods of air filtration.
8.9.1 Straining. Straining occurs when a particle is larger
than the opening between fibers and cannot pass through. It is
a very ineffective method of filtration because the vast majority of particles are far smaller than the spaces between fibers.
Straining will remove lint, hair, and other large particles.
The wide range in performance of in-line media-style air filters made it necessary to agree on a new consolidated method
of efficiency testing. The new adopted, industry-accepted
method in the United States is the minimum efficiency reporting value (MERV) system developed by ASHRAE. This filter
rating system ranges from 1 through 20, where a rating of 1 is
a very coarse, see-through style home HVAC filter and a rating
of 20 exceeds even the ability of a HEPA (High Efficiency
Particulate Air) filter. In a HEPA dioctylphthalate (DOP) Test,
0.3 micron particles of dioctylphthalate (DOP) are drawn
through a HEPA filter. Efficiency is determined by comparing
the downstream and upstream particle counts. To be designated as a HEPA filter, the filter should be at least 99.97% efficient, i.e., only three particles of 0.3 micron size can pass for
every 10,000 particles fed to the filter.
1.5–3.5 [375–875]
2.5–6 [625–1500]
Note 2
2.5–11 [625–2750]
1–5
1–5
1–2
1–5
3.0–6.0 [760–1480]
3–12 [750–3000]
Note 2
20–40
10–30
10–20
—
—
—
2–4 [500–1000]
5–10 [650–1300]
10–100 [2500–25000] 5–15 [650–1950]
5–10 [650–1300]
3–5 [390–650]
0.5–1 [65–130]
10–40 [1300–5200]
Large
Moderate
Small
Moderate
Moderate
Large
Moderate
Small
Small
Large
Large
Large
Moderate
Large
Large
—
—
—
—
—
—
Space
0.5–5
0.5–2
{
3–6
3–6
3–6
3–6
3–8
0.25
0.25
0.25
0.25
0.25
[750–1500]
[750–1500]
[750–1500]
[750–1500]
[750–2000]
0.5 [125]
0.25
Pressure
Loss inches [Pa]
H2O Gal. Per
1000 acfm
[1000 L/am3 – gas]
Negligible
Negligible
Negligible
Negligible
Negligible
Yes
Yes
No
Varies with
Design
%Q
%Q
%Q
%Q
%Q
%Q
% (Q)2
Yes
Yes
No
% (Q)2
% (Q)2
Note 2
Slightly
Yes
% (Q)2
% (Q)2
% Q or less
Note 2
Yes
Efficiency
Negligible
Sensitivity to Q Change
Pressure
Note 1: Pressure loss is that for fabric and dust cake. Pressure losses associated with outlet connections to be added by system designer.
Note 2: A function of the mechanical efficiency of these combined exhausters and dust collectors.
Note 3: Precooling of high temperature gases will be necessary to prevent rapid evaporation of fine droplets.
Note 4: See NFPA requirements for fire hazards, e.g., zirconium, magnesium, aluminum, woodworking, etc.
Higher Efficiency:
Fog Tower
Venturi
Dry Centrifugal:
Low Pressure Cyclone
High Eff. Centrifugal
Dry Dynamic
Electrostatic:
Fabric:
Intermittent—Shaker
Continuous—Shaker
Continuous—Reverse Air
Continuous—Reverse Pulse
Glass, Reverse Flow
Wet:
Packed Tower
Wet Centrifugal
Wet Dynamic
Orifice Types
Type
Higher Efficiency
Range on Particles
Greater than Mean
Size in Microns
TABLE 8-4. Comparison of Some Important Dust Collector Characteristics
Note 1
{
400 [204]
400 [204]
Note 3
Unlimited
%- Proportional to
May cause
condensation
and plugging
None
500 [260]
See Table 8-1
Unlimited
{
{ {
None
May make
reconditioning
difficult
Improves efficiency 500 [260]
Max. Temp. F [C]
Standard
Construction
Humid Air Influence
Note 4
8-38
Industrial Ventilation
Air Cleaning Devices
TABLE 8-5. Media Velocity vs. Fiber Size
8.10
Filter Type
Filter Size
(microns)
Velocity
fpm [m/s]
Media
Filtration
Mechanism
Panel Filters
25–50
250–625
Impingement
25–50
500
Impingement
[2.50]
Extended Surface Filters
0.75–2.5
20–25
Interception
[0.10–0.13]
HEPA Filters
0.5–6.3
5
RADIOACTIVE AND HIGH TOXICITY OPERATIONS
There are three major requirements for air cleaning equipment to be utilized for radioactive or high toxicity applications:
1) High efficiency,
2) Low maintenance, and
[1.25–3.20]
Automatic Roll Filters
8-39
Diffusion
[0.25]
MERV filters come in four typical filter types, as follows:
Flat or panel air filters with a MERV of 1 to 4 are commonly
used in residential furnaces and air conditioners. They are
NOT typically used in industrial ventilation applications.
Second, there are pleated or extended surface filters, with a
MERV of 5 to 15 range from 1" [25 mm] deep pleated filters
to true box and envelope filters. Third are high efficiency box
and envelope filters, with a MERV of 14 to 16. Finally, there
are true HEPA filters (MERV 17 to 20). Figure 8-19 shows the
general relationship. Table 8-6 compares several important
characteristics of commonly used air filters. Considerable life
extension of an expensive final filter can be obtained by the
use of one or more cheaper, less efficient, prefilters. For example, the life of a HEPA filter can be increased 25% with a
“throwaway” prefilter. If the “throwaway” filter is followed by
a 90% efficient extended surface filter, the life of the HEPA filter can be extended nearly 900%. This concept of progressive
filtration allows the final filters in clean rooms to remain in
place for 10 years or more.
3) Safe disposal.
High efficiency is essential because of extremely low tolerances for the quantity and concentration of stack effluent and
the high cost of the materials handled. Not only should the efficiency be high, it should also be verifiable because of the legal
requirement to account for all radioactive material.
The need for low maintenance is of special importance
when exhausting any hazardous material. For many radioactive processes, the changing of bags in a conventional fabric
collector may expend the daily radiation tolerances of 20 or
more persons. Infrequent, simple, and rapid maintenance
requirements are vital. Another important factor is the desirability of low residual build up of material in the collector
since dose rates increase with the amount of material and
reduce the allowable working time.
Disposal of radioactive or toxic materials is a serious and
difficult problem. For example, scalping filters loaded with
radioactive dust are usually incinerated to reduce the quantity
of material that should be disposed of in special burial
grounds. The incinerator will require an air cleaning device,
such as a wet collector of very special design, to avoid unacceptable pollution of ambient air and water.
With these factors involved, it is necessary to select an air
cleaning device that will meet efficiency requirements without
causing difficulty in handling and disposal.
Cartridge-style, dust collector filter units especially
designed for high efficiency and low maintenance are available. These units feature quick changeout through a plastic
barrier (bag-in, bag-out), which is intended to encapsulate
FIGURE 8-19. Comparison between various methods of measuring air cleaning capability
8-40
Industrial Ventilation
TABLE 8-6. Comparison of Some Important Air Filter Characteristics
Pressure Drop "wg [Pa]
(Notes 1 & 2)
ASHRAE Performance
(Note 4)
Maintenance
(Note 6)
Initial
Final
MERV
(Note 5)
Arrestance
Efficiency
1. Glass Throwaway
(2" deep)
0.1 [25]
0.5 [125]
2–3
77%
2. High Velocity
(permanent units)
(2" deep)
0.1 [25]
0.5 [125]
2–3
3. Automatic
(viscous)
0.4 [100]
0.4 [100]
0.15–0.60
[40–150]
Face Velocity
fpm [m/s]
Labor
Material
NA
Note 7
300 [1.50]
High
High
73%
NA
Note 7
500 [2.5]
High
Low
3
80%
NA
Note 7
500 [2.5]
Low
Low
0.5–1.25
[125–310]
8–12
90–99%
25–95%
300–625
[1.50–3.2]
Medium
Medium
a. Dry Agglomerator/ 0.35 [90]
Roll Media
0.35 [90]
10–12
NA
Note 8
90%
500 [2.50]
Medium
Low
b. Dry Agglomerator/ 0.55 [140]
Extended Surface
Media
1.25 [310]
13–16
NA
Note 8
95%+
530 [2.70]
Medium
Medium
c. Automatic Wash
Type
0.25 [60]
0.25 [60]
13–16
NA
Note 8
95.5
400–600
[2.00–3.00]
Low
Low
0.5–1.0
[125–250]
1.0–3.0
[250–750]
17–20
Note 3
Note 3
250–500
[1.25–2.50]
High
High
Type
Low/Medium Efficiency
Medium/High Efficiency
1. Extended Surface
(dry)
2. Electrostatic
Ultra High Efficiency
1. HEPA
Note 1: Pressure drop values shown constitute a range or average, whichever is applicable.
Note 2: Final pressure drop indicates point at which filter or filter media is removed and the media is either cleaned or replaced. All others are cleaned in place, automatically,
manually, or media renewed automatically. Therefore, pressure drop remains approximately constant.
Note 3: 95–99.97% by particle count, DOP test.
Note 4: ASHRAE Standard 52-76 defines (a) Arrestance as a measure of the ability to remove injected synthetic dust, calculated as a percentage on a weight basis and (b)
Efficiency as a measure of the ability to remove atmospheric dust determined on a light-transmission (dust spot) basis.
Note 5: ASHRAE MERV (Minimum Efficiency Reporting Value) Efficiencies range from 1 (lowest) through 20 (highest).
Note 6: Compared to other types within efficiency category.
Note 7: Too low to be meaningful.
Note 8: Too high to be meaningful.
spent filters, thereby eliminating the exposure of personnel to
radioactive or toxic material.
EXPLOSION VENTING/DEFLAGRATION VENTING
most organic dusts such as grain, wood, plastics, coal and
many others. Metal dust is also prone to deflagrations and can
be especially dangerous. Many of these dusts and associated
industries have their own National Fire Protection Association (NFPA) designated Code (see reference 8.11).
Two distinct types of explosions exist in nature. A detonation is an explosion that propagates at a velocity in excess of
the speed of sound and cannot be controlled. In a deflagration,
the combustion wave propagates more slowly (at less than the
speed of sound) and can be controlled, if designed properly.
Examples of detonations include dynamite, solid rocket fuel
or other similar material. Examples of deflagrations include
To begin taking precautions, sources of possible ignition
should be identified and controlled to minimize the risk of a
dust cloud explosion. Usual causes of explosions include static
discharge, hot surfaces on machinery and sparks and flames
from processes. After identifying possible sources of ignition,
preventive measures should be taken. Static grounding of the
equipment and spark traps are typical preventive measures.
For further information on this subject, see reference 8.10.
8.11
Air Cleaning Devices
The addition of an inert gas to replace oxygen in a dust collector can prevent an explosion by ensuring the minimum oxygen
content required for ignition is never reached. Inerting can be
very effective in closed loop systems but is not economical in
typical local exhaust systems because of the constant loss of
expensive inerting gas. Should ignition occur, protective measures should be taken to limit the damage. Typical protective
measures include: explosion suppression, explosion containment, and explosion venting.
Explosion suppression requires the early detection of an
explosion, usually within the first 20 milliseconds. Once ignition is detected, an explosion suppression device injects a pressurized chemical suppressant into the collector to displace the
oxygen or inert the combustible particles and impede combustion. These systems can be very useful when toxic dusts are
being handled.
Explosion containment uses specialized dust collectors
designed to withstand the maximum pressure generated to
contain the explosion. Most pressure capabilities of commercially available dust collectors are not designed sufficiently to
contain an explosion in progress. Collection equipment
designed for containment often requires ASME pressure vessel code construction. A competent manufacturer should be
consulted for the requirements of this type of equipment.
Explosion venting, the most common protection, is afforded by fitting pressure relief vents to the collector housing.
Standard vents can only be used when the dust collector is
located outside or immediately adjacent to an outdoor wall. If
the collector is located indoors and is not near an exterior
wall, a flameless vent(s) may be a possibility. As pressure
increases quickly leading up to an explosion, a relief vent
opens to allow the rapidly expanding gases to escape. This
effectively limits the maximum pressure build up to less than
the bursting pressure of the vessel. The necessary area for
such a relief vent is a function of the vessel volume, vessel
height, vessel strength, the opening pressure of the relief
vent, and the rate of pressure rise characteristic of the dust in
question. Most standard dust collectors will require reinforcing to withstand the reduced maximum pressure experienced
during an explosion.
To choose the most reliable, economical, and effective means
of explosion control, an evaluation of the specifics of the exhaust
system and the degree of protection required is necessary.
The NFPA Standards(8.11) are the most commonly recognized
standards and should be studied and thoroughly familiar to anyone responsible for the design or evaluation of dust collectors
applied to potentially explosive dusts. Always verify that the latest issue is being referenced for data and standards. In addition,
system designers should cooperate with authorities having jurisdiction (AHJs) for design consensus and compliance. Additional
information regarding considerations for NFPA combustible
dust system design is provided in Chapter 12.
8-41
REFERENCES:
8.1
Leith, D.; First, M.K.W.; Feldman, H.: Performance of
a Pulse-Jet at High Velocity Filtration II, Filter Cake
Redeposition. J. Air Pollut. Control Assoc. 28:696
(July 1978).
8.2
Beake, E.: Optimizing Filtration Parameters. J. Air
Pollut. Control Assoc. 24:1150 (1974).
8.3
Leith, D.; Gibson, D. D.; First, M. W.: Performance of
Top and Bottom Inlet Pulse-Jet Fabric Filters. J. Air
Pollut. Control Assoc. 24:1150 (1974).
8.4
American Society of Heating, Refrigerating and AirConditioning Engineers, Inc.: HVAC Systems and
Equipment Handbook. Atlanta, GA (1996).
8.5
Lund, H.F.: Industrial Pollution Control Handbook.
McGraw-Hill (1971).
8.6
Heumann, W.L.: Industrial Air Pollution Control
Systems. McGraw-Hill (1997).
8.7
Gilliland, G.A.; Ramaswami, R.D.; Patel, D.N.:
Removal of Volatile Organic Compounds (VOCs)
Generated by Forest Product Industries Using
Biofiltration Technology. In Proc. Emerging
Technologies in Hazardous Waste Management VII,
ACS Special Symposium: Atlanta, GA, September,
17-20, 1995. Tedder, D.W., Editor, Washington, DC
(United States) American Chemical Society p. 921
(1352p) CONF-9509139.
8.8
Biofiltration. Air emissions from Wood and WoodBased Products: Conducting Research and Sharing
Information. 22 April 1998. USDA Forest Products
Laboratory. 16 Dec 2000. http.fpl.fs.fed.us/voc/
biofilt.html.
8.9
American Society of Heating, Refrigerating and AirConditioning Engineers: Method of Testing Cleaning
Devices Used in General Ventilation for Removing
Particulate Matter. ASHRAE Pub. No. 52-76.
ASHRAE, Atlanta, GA (May 1976).
8.10
National Council on Radiation Protection and
Measurement: NCRP Report No. 39, Basic Radiation
Protection Criteria. NCRP Report No. 39.
Publications, Bethesda, MD (January, 1971).
8.11
NFPA 654: Standard for the Prevention of Fire and
Dust Explosions from the Manufacturing, Processing,
and Handling of Combustible Particulate Solids
(2006); NFPA 68: Guide for Venting of Deflagrations
(2014); NFPA 69: Standard on Explosion Prevention
Systems (2002); NFPA 91: Standard for Exhaust
Systems for Air Conveying of Vapors, Gases, Mists,
and Noncombustible Particulate Solids (2004); NFPA
484: Standard for Combustible Metals (2015); NFPA
497: Recommended Practice for the Classification of
Flammable Liquids, Gases, or Vapors and of
Hazardous (Classified) Locations for Electrical
8-42
Industrial Ventilation
Installations in Chemical Process Areas (2004),
National Fire Protection Association, Quincy, MA.
NFPA 33: Standard for Spray Applications Using
Flammable or Combustible Materials; NFPA 61:
Standard for the Prevention of Fires and Dust
Explosions in Agricultural and Food Processing
Facilities; NFPA 61: Standards for the Prevention of
Fires and Dust Explosions in Agricultural and Food
Processing Facilities; NFPA 652: Standards on
Fundamentals of Combustible Dust (2015); NFPA
664: Standard for the Prevention of Fires and
Explosions in Wood Processing and Woodworking
Facilities (2012).
8.12
Farr Company, El Segundo, CA (May, 2011).
8.13
Duall Division of MetPro/Ceco Environmental.
Owosso, MI (March, 2010).
8.14
Aget Mfg. Company, Adrian, MI (June, 2010).
8.15
J. Kirt Boston. Minneapolis, MN (June, 2014).
8.16
C.T. Womack. Statesville, NC (June, 2014).
8.17
Dan Josephs, Louisville, KY (July, 2014).
In SI units, the same problem would be set up as follows:
36,000 acfm Y 16.9884 am3/s
120 F Y 49 C
@ 100% humidity Y 80 g/kg dry air (Chapter 9,
Figure 9-m)
Humid Volume Y 1.03 m3/kg dry air
APPENDIX A8 CONVERSION OF POUNDS PER HOUR
(EMISSIONS RATE) TO GRAINS PER DRY STANDARD
CUBIC FOOT [g/nm3 (dry)] (EMISSION DENSITY OR
LOADING)
A collector is measured with a flow of 36,000 acfm of air at
120 F, 100% humidity and a particulate mass emissions rate of
1 pound per hour. What is the emissions rate in terms of grains
per dry standard cubic foot (gr/dscf)?
120 F dB, 100% humidity 6 0.0816 pounds H2O/pound of
dry air = 571 grains of water/pound dry air (Chapter 9, Figure
9-i)
Humid volume = 16.56 ft3 per pound of dry air
1 pound particulate Y 453.6 grams
453 grams/hr × 1 hr/3600 s = 0.1258 grams/s
Chapter 9
LOCAL EXHAUST VENTILATION SYSTEM DESIGN
CALCULATION PROCEDURES
NOTE: Equations with notation followed by (IP) are designated for inch-pound system only; equations followed by (SI) are designated for
metric use only. If equation bears neither, then it applies to both systems.
9.1
9.2
9.3
9.4
9.5
9.6
INTRODUCTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-3
PRELIMINARY SYSTEM DESIGN
INFORMATION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-4
DESIGN CONSIDERATIONS FOR
CALCULATING SYSTEM AIRFLOW RATES
AND RESISTANCE LOSSES . . . . . . . . . . . . . . . . . . . .9-4
9.3.1 Airflow Rate (Q) . . . . . . . . . . . . . . . . . . . . . . . .9-4
9.3.2 Determining Resistance Losses – System
Component Loss Factors . . . . . . . . . . . . . . . . . .9-4
9.3.3 Friction Loss in Round Straight Duct . . . . . . . .9-5
9.3.4 Friction Loss in Non-Circular Straight Duct . . .9-5
9.3.5 Friction Loss in Flexible Straight Duct . . . . . . .9-5
9.3.6 Dynamic Losses in Elbows . . . . . . . . . . . . . . . .9-7
9.3.7 Dynamic Losses in Branch Entries . . . . . . . . . .9-8
STATIC PRESSURE LOSSES – SPECIAL
CONSIDERATIONS . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-8
9.4.1 Contractions and Expansions . . . . . . . . . . . . . . .9-8
9.4.2 Special Expansion Considerations – Evasé
Discharge . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-8
9.4.3 Determining the Loss in Traps and Settling
Chambers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-8
BASIC SYSTEM DESIGN PROCEDURES AND
CALCULATIONS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-8
9.5.1 Hood Airflow at Nonstandard (Actual)
Conditions . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-11
9.5.2 Addition of Materials Inside the Hood . . . . . .9-12
9.5.3 Combining of Gases of Different Densities
Due to Temperature . . . . . . . . . . . . . . . . . . . . .9-13
CALCULATION SHEET DESIGN PROCEDURE . . .9-14
9.6.1 Use of the Velocity Pressure Method . . . . . . .9-14
9.6.2 Use of the Calculation Sheet . . . . . . . . . . . . . .9-14
9.6.3 Calculation and Input of System Design
Data on the Calculation Sheet . . . . . . . . . . . . .9-15
9.7
SAMPLE SYSTEM DESIGN #1 (SINGLEBRANCH SYSTEM AT STANDARD AIR
CONDITIONS) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-16
9.8 DISTRIBUTION OF AIRFLOW IN A MULTIBRANCH DUCT SYSTEM . . . . . . . . . . . . . . . . . . . . .9-22
9.8.1 Use of the Balance-by-Design (Static
Pressure Balance) Method . . . . . . . . . . . . . . . .9-23
9.8.2 Use of the Blast Gate/Orifice Plate Method . .9-23
9.9 INCREASING VELOCITY THROUGH A
JUNCTION (WEIGHTED AVERAGE
VELOCITY PRESSURE) . . . . . . . . . . . . . . . . . . . . . . . . .9-24
9.10 SYSTEM AND FAN PRESSURE CALCULATIONS . .9-25
9.10.1 System Static Pressure (SSP) . . . . . . . . . . . . . .9-25
9.10.2 Fan Total Pressure (FTP) . . . . . . . . . . . . . . . . .9-25
9.10.3 Fan Static Pressure (FSP) . . . . . . . . . . . . . . . .9-25
9.10.4 Use of System Static Pressure to
Specify a Fan . . . . . . . . . . . . . . . . . . . . . . . . . .9-25
9.11 THE SYSTEM AND FAN CURVE RELATIONSHIP . .9-26
9.12 SAMPLE SYSTEM DESIGN #2 (MULTI-BRANCH
SYSTEM AT STANDARD AIR CONDITIONS) . . . .9-27
9.13 CALCULATION METHODS AND NONSTANDARD AIR DENSITY . . . . . . . . . . . . . . . . . . . .9-33
9.14 SAMPLE SYSTEM DESIGN #3 (SINGLEBRANCH SYSTEM AT NON-STANDARD
AIR CONDITIONS) (IP UNITS ONLY) . . . . . . . . . . .9-33
9.15 SAMPLE SYSTEM DESIGN #4 (ADDING
A BRANCH TO AN EXISTING SYSTEM
AT NON-STANDARD AIR CONDITIONS)
(IP UNITS ONLY) . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-38
9.16 AIR BLEED DESIGN . . . . . . . . . . . . . . . . . . . . . . . . . .9-41
REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-42
APPENDIX A9 PRESSURE MEASUREMENT
IN THE SI SYSTEM . . . . . . . . . . . . . . . . . . . . . . . . . . .9-42
____________________________________________________________
Figure 9-1
Figure 9-2
Fitting and Duct Losses . . . . . . . . . . . . . . . . . .9-6
System Duct Calculation Parameter
Location . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-7
Figure 9-3 (IP) Expansions and Contractions . . . . . . . . . . . . . .9-9
Figure 9-3 (SI) Expansions and Contractions . . . . . . . . . . . . .9-10
Figure 9-4
Data Entry to Calculation Sheet
(Example Problem 9-7) . . . . . . . . . . . . . . . . .9-17
Figure 9-5 (IP) Double Line Sketch (Sample System
Design #1 . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-18
Figure 9-5 (SI) Double Line Sketch (Sample System
Design #1 . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-19
Figure 9-6 (IP) Velocity Pressure Method
Calculation Sheet . . . . . . . . . . . . . . . . . . . . . .9-20
9-2
Industrial Ventilation
Figure 9-6 (SI) Velocity Pressure Method
Calculation Sheet . . . . . . . . . . . . . . . . . . . . . .9-21
Figure 9-7 (IP) Branch Entry Velocity Correction . . . . . . . . .9-25
Figure 9-7 (SI) Branch Entry Velocity Correction . . . . . . . . .9-25
Figure 9-8
Sample Bulk Powder Handling System –
Sample System Design #2 . . . . . . . . . . . . . . .9-27
Figure 9-9
Single Line Sketch – Sample System
Design #2 . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-28
Figure 9-10
Elevation Drawing – Sample System
Design #2 . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-28
Figure 9-11
Basic System Information – Sample
System Design #2 . . . . . . . . . . . . . . . . . . . . .9-29
Figure 9-12 (IP) Velocity Pressure Method Calculation
Sheet – Sample System Design #2 . . . . . . . .9-30
Figure 9-12 (SI) Velocity Pressure Method Calculation
Sheet – Sample System Design #2 . . . . . . . .9-31
Figure 9-13
System Layout – Sample System
Design #3 . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-34
Figure 9-14(IP) Velocity Pressure Method Calculation
Sheet – Sample System Design #3 . . . . . . . .9-35
Figure 9-15
Fan Rating Table . . . . . . . . . . . . . . . . . . . . . .9-36
Figure 9-16
Psychrometric Chart for Humid Air
(excerpted from Figure 9-j (IP)) . . . . . . . . . .9-37
Figure 9-17
System Layout (Sample System
Design #4) . . . . . . . . . . . . . . . . . . . . . . . . . . .9-39
Figure 9-18(IP) Velocity Pressure Method Calculation
Sheet – Sample System Design #4 . . . . . . . .9-40
Figure 9-19
Air Bleed Opening . . . . . . . . . . . . . . . . . . . . .9-41
Design Factors and Charts
Figure 9-a
Hood Entry Loss Factors . . . . . . . . . . . . . . . .9-55
Figure 9-b(IP) Friction Chart for Sheet Metal &
Plastic Ducts . . . . . . . . . . . . . . . . . . . . . . . . . .9-56
Figure 9-c(IP) Friction Chart for Sheet Metal &
Plastic Ducts . . . . . . . . . . . . . . . . . . . . . . . . . .9-57
Figure 9-d
Expansions and Contractions . . . . . . . . . . . . .9-58
Figure 9-e
Duct Design Data Elbow Loss Factors . . . . .9-59
Figure 9-f
Branch Entry Loss Factors and Losses
in Settling Chambers . . . . . . . . . . . . . . . . . . .9-60
Figure 9-g
Weather Cap Losses . . . . . . . . . . . . . . . . . . . .9-61
Figure 9-h (IP) Psychrometric Chart – 30 F to 115 F DB
Temperature . . . . . . . . . . . . . . . . . . . . . . . . . .9-62
Figure 9-i (IP) Psychrometric Chart – 60 F to 250 F DB
Temperature . . . . . . . . . . . . . . . . . . . . . . . . . .9-63
Figure 9-j(IP) Psychrometric Chart – 100 F to 500 F DB
Temperature . . . . . . . . . . . . . . . . . . . . . . . . . .9-64
Figure 9-k (IP) Psychrometric Chart – Up to 1500 F DB
Temperature . . . . . . . . . . . . . . . . . . . . . . . . . .9-65
Figure 9-l (SI) Psychrometric Chart – 0 C to 50 C DB
Temperature . . . . . . . . . . . . . . . . . . . . . . . . . .9-66
Figure 9-m (SI) Psychrometric Chart – 10 C to 120 C DB
Temperature . . . . . . . . . . . . . . . . . . . . . . . . . .9-67
Figure 9-n (SI) Psychrometric Chart – 100 C to 200 C DB
Temperature . . . . . . . . . . . . . . . . . . . . . . . . . .9-68
____________________________________________________________
Table 9-1
Table 9-2
Table 9-3 (IP)
Table 9-3 (SI)
Table 9-4 (IP)
Table 9-4 (SI)
Table 9-5(IP)
Abbreviations Used in Chapter . . . . . . . . . . . .9-3
Area and Circumference of Circles . . . . . . . .9-43
Velocity Pressure to Velocity Conversion
– Standard Air . . . . . . . . . . . . . . . . . . . . . . . .9-44
Velocity Pressure to Velocity Conversion
– Standard Air . . . . . . . . . . . . . . . . . . . . . . . .9-45
Velocity to Velocity Pressure Conversion
– Standard Air . . . . . . . . . . . . . . . . . . . . . . . .9-46
Velocity to Velocity Pressure Conversion
– Standard Air . . . . . . . . . . . . . . . . . . . . . . . .9-47
Duct Friction Loss Factors per Foot of
Duct Length, F'd . . . . . . . . . . . . . . . . . . . . . . .9-48
Table 9-5 (SI)
Table 9-6 (IP)
Table 9-6 (SI)
Table 9-7 (IP)
Table 9-7 (SI)
Duct Friction Loss Factors per Meter of
Duct Length, F'd . . . . . . . . . . . . . . . . . . . . . . .9-50
Circular Equivalents of Rectangular
Ducts (in) . . . . . . . . . . . . . . . . . . . . . . . . . . . .9-52
Circular Equivalents of Rectangular
Ducts (mm) . . . . . . . . . . . . . . . . . . . . . . . . . .9-53
Air Density Correction Factor (Temperature and Elevation Only), dfT H dfe . . . . . . .9-54
Air Density Correction Factor (Temperature and Elevation Only), dfT H dfe . . . . . . .9-54
Local Exhaust Ventilation System Design Calculation Procedures
9-3
TABLE 9-1. Abbreviations Used in Chapter
= inches water gauge (pressure unit in IP system)
"wg
L
= length
A
2 [m2]
= area in ft
LEV
= local exhaust ventilation
acfm
= actual cubic feet per minute
m
= meters
am3/s
= actual cubic meters per second
ṁ
= mass flow rate in lbm/min [kg/s]
ASL
= above sea level
min
= minutes
BHP
= brake horsepower
mm
= millimeters
C
= degrees Celsius
P
= pressure in "wg [Pa]
CP
= heat capacity in BTU/lbm-R [kJ/kg-K]
Pa
= Pascals
d
= diameter
Pa
= actual pressure in "wg [Pa]
dequiv
= equivalent diameter
Pe
= equivalent pressure in "wg [Pa]
da
= dry air
Q
= airflow rate
df
= density factor
Qact
= actual airflow rate in acfm [am3/s]
dfe
= density factor for elevation
Qcorr
= corrected airflow rate in acfm [am3/s]
dfm
= density factor for moisture
Qstd
= standard airflow rate in dscfm [dnm3/s]
dfP
= density factor for pressure
r
= radius
dft
= density factor for temperature
Rg
= specific gas constant
dnm3/s
= dry normal cubic meters per second
RPM
= revolutions per minute
dscfm = dry standard cubic feet per minute
s
= seconds
F
= degrees Fahrenheit
SP
= static pressure in "wg [Pa]
fD
= Darcy friction factor
SPgov
= governing static pressure in "wg [Pa]
= static pressure into the fan in "wg [Pa]
ft
= feet
SPin
Fcont
= contraction loss factor
SPlower = lower static pressure in "wg [Pa]
Fd
= duct friction loss factor
SPout
= static pressure out of the fan in "wg [Pa]
F′d
= duct friction loss factor per foot [meter]
SSP
= system static pressure in "wg [Pa]
Fel
= elbow (90-degree) loss factor
T
= temperature in F [C] (or R [K])
Fen
= branch entry loss factor
TP
= total pressure in "wg [Pa]
Fexp
= expansion regain factor
V
= velocity in fpm [m/s]
FSP
= fan static pressure in "wg [Pa]
VP
= velocity pressure in "wg [Pa]
FTP
= fan total pressure in "wg [Pa]
VPr
= weighted average velocity pressure in "wg [Pa]
h
= enthalpy (total energy) in BTU/lbm [kJ/kg]
VPd
= duct velocity pressure in "wg [Pa]
H
= height
VPin
= velocity pressure into the fan in "wg [Pa]
HP
= horsepower
VPout
= velocity pressure out of fan in "wg [Pa]
kg
= kilograms
W
= width
lbm
= pounds mass

= moisture content in lbm H2O/lbm da [kg H2O/kg da]
IVS
= industrial ventilation system
z
= elevation in feet [meters]
K
= degrees Kelvin
9.1
INTRODUCTION
Most local exhaust ventilation (LEV) systems are comprised of a combination of hoods, duct components, an air
cleaning device(s), a fan and a stack. To ensure proper performance and achieve economic efficiency, such systems must
be properly designed, balanced and commissioned. This
process involves more than simply connecting system components together. If LEV systems are not designed in a manner
that ensures that all design flow rates will be realized, contaminant control may not be achieved in an economically desirable manner.
Additionally, failure to ensure proper design may result in
9-4
Industrial Ventilation
be controlled. This includes the elevation, temperature,
pressure, moisture content and heat content of the
airstream for each process and duct segment. The correct determination and use of the density factor, df, for
air/gas streams (used interchangeably in this Manual)
are crucial to performing accurate LEV system design
calculations. See Chapter 3, Section 3.5,
the settling of particulate contaminants inside a duct system
when minimum duct transport velocities are not maintained.
This condition is detrimental to system performance and poses
a serious hazard to employees; particularly if the dust is combustible.
Proper design of an LEV system requires a thorough understanding of the following chapters of this Manual:
•
Chapter 3: Principles of Airflow
•
Chapter 4: Industrial Ventilation System Design
Principles
•
Chapter 5: Duct System and Discharge Stack Design
Principles
•
Chapter 6: Hood Design
•
Chapter 7: Fans
•
Chapter 8: Air Cleaning Devices
Failure to properly grasp the material presented in the identified chapters may result in the design and installation of an
LEV system that does not achieve the desired goals. With such
an understanding, and by following the design calculation procedure presented in this chapter, the proper specification of
hood airflows, determination of appropriate duct sizes, and
computation of the System Static Pressure (SSP) may be economically achieved.
To ensure the proper design of LEV systems, the density of
the airstreams moved through the system must be considered.
Density is impacted by elevation, temperature, pressure and
moisture content. Reference Chapter 3 for a more thorough
discussion and understanding of the density of airstreams.
9.2
PRELIMINARY SYSTEM DESIGN INFORMATION
Chapter 4 details information that should be gathered prior to
beginning the system design process. This information includes:
•
A layout of the operations, workroom, and building (if
necessary),
•
The available location(s) for the air cleaning device(s)
and the fan(s),
•
A line sketch of the duct system layout, including plan
and elevation dimensions and the location of the air
cleaning device(s), fan(s), and other pertinent components. For convenience, number, letter, or otherwise
identify each hood, branch, and section of main duct.
(The examples herein show hoods numbered and other
system segment points lettered.),
•
A design or sketch of the desired hood for each operation with the direction and elevation of the outlet for
each duct connection,
•
Specific details about required capture velocities, flow
rates, hood entry losses and minimum transport velocities,
•
Information regarding the density of the airstream(s) to
•
9.3
The method and location of replacement air distribution devices. Such devices impact hood performance.
The type and location of these fixtures can dramatically
decrease contaminant control by creating undesirable
turbulence at the hood face (see Chapter 11).
DESIGN CONSIDERATIONS FOR CALCULATING
SYSTEM AIRFLOW RATES AND RESISTANCE
LOSSES
Successful LEV system design requires the performance of
a series of calculations to properly size and specify all system
components. Two primary system characteristics that must be
known to successfully design an industrial ventilation system
(IVS) are airflow rate and resistance (i.e., static pressure) losses for the various system components. The procedure established in this chapter permits the designer to size each component and determine the airflow and resistance in each segment
of the system. This information will then be used to specify
key pieces of equipment such as the fan and air cleaning
device.
9.3.1 Airflow Rate (Q). Sometimes called volume or airflow, the appropriate airflow rate for each hood is determined
either by formulae for specific hood designs or by use of empirical data and experience specific processes. Chapter 13 identifies empirically derived airflow rates for many specific hood
designs. In cases where such empirical information is not available, Chapter 6 identifies considerations and equations for calculating hood airflows. In all cases, airflow rates are in acfm
[am3/s] at local conditions. For example, the volume shown in
VS-15-02 is listed as 400 to 500 acfm [0.20 to 0.25 am3/s] for
non-toxic dust. Given that empirical and calculated flow rates
are denoted at actual conditions, a thorough understanding of
the characteristics of an airstream is pertinent to proper LEV
system design (see Chapter 3, Sections 3.3 through 3.5).
9.3.2 Determining Resistance Losses – System Component Loss Factors. Once the proper airflow has been deter-
mined, next calculate resistances to the flow of that air through
the LEV system. The sum of these resistances is known as the
system static pressure (SSP), that is used for fan selection purposes. SSP in the inch-pound (IP) system is commonly measured in units of inches-water gauge ("wg); in the International
System of Measure (SI), it is commonly measured in pascals
(Pa). See Appendix B for conversion factors to convert other
units of measure for pressure.
The resistances in a system that combine to create its SSP
are derived from a loss factor multiplied by an appropriate
Local Exhaust Ventilation System Design Calculation Procedures
velocity pressure (VP) (see Chapter 3, Figures 9-1 and 9-2,
and Tables 9-3 and 9-4). The procedure used to determine the
SSP is known as the velocity pressure method, or VP method.
These loss factors, which represent frictional and/or dynamic
losses in the system, are expressed as multipliers of VP. Values
for these loss factors are acquired from laboratory and mathematical methods as well as empirical evidence. Loss factors
are specified or derived from information contained in
Chapters 3, 6 and 13, Table 9-5, and Figures 9-a through 9-g.
For convenience, loss factors for commonly used elbows
and branch entries are located on the right edge of the ACGIH®
Calculation Sheet (Figure 9-6; called the calc sheet herein) discussed in Section 9.7. Figure 9-2 shows the location and application of these various loss factors in a simple hood and duct
branch segment.
9.3.3 Friction Loss in Round Straight Duct. The resistance to airflow due to friction in the duct is a function of the
duct friction loss factor (Fd) multiplied by the duct velocity
pressure (VPd). The duct friction loss factor is a product of the
duct friction loss factor per foot [meter] of duct (F′d) multiplied
by the length of straight duct (L). Note that duct segments containing fittings (e.g., elbows, branch entries, expansions and
contractions) must be measured to the centerline of the fitting
to properly account for the friction loss experienced in the fittings.
The duct friction loss and duct friction loss factor are
expressed as:
Duct Friction Loss = (Fd)(VP)
[9.1]
Fd = (F′d)(L)
[9.2]
The duct friction loss factor per foot for all metal and plastic
duct is expressed as:
F′d(met./plas.) = 0.0307(V0.533/Q0.612)
[9.3] IP
In SI units, the duct friction loss factor per meter for all metal
and plastic duct is:
F′d(met./plas.) = 0.0155(V0.533/Q0.612)
[9.4] SI
Duct friction loss factors per foot [meter] of duct are also
presented in this chapter in Table 9-5 (IP and SI) as well as in
Figures 9-b and 9-c.
The LEV system design process was simplified to use one
set of factors for all metal and plastic ducts since duct will be
uniformly coated with dust and other materials after some
period of operation. Note that a different equation must still be
used to determine the duct friction loss factor per foot [meter]
for flexible ducts.
NOTE: Unless specified otherwise, the use of sheet metal
duct is assumed in all problems throughout this chapter.
The duct friction loss factor per foot [meter] equations (IP
and SI) noted above are also located on the right side of the
calc sheet. Values for F′d are also presented in Table 9-5 (IP
9-5
and SI). Table 9-5 lists the factors as a function of duct diameter for six different velocities; linear interpolation between
velocity values may be performed. Additionally, these values
may be selected from information in Figures 9-b and 9-c.
9.3.4 Friction Loss in Non-Circular Straight Duct. Round
ducts are strongly recommended for use with LEV systems.
They produce a more uniform air velocity profile that resists
settling of particulate material and are capable of withstanding
higher static pressures. At times, however, the designer must
use other duct shapes.
To determine the friction loss in a rectangular duct determine the duct’s equivalent diameter, dequiv, by using Table 9-6
or Equation 9.5. Once the rectangular duct’s equivalent diameter has been determined, the same process for determining the
duct friction loss factor per foot [meter] for round duct is then
followed on the basis of equal friction loss.
It is critical to maintain the minimum transport velocity in
rectangular duct. However, even if the minimum transport
velocity requirement is met, the flow characteristics in rectangular ducts could yield dead spots and potential locations for
material to settle out in corners. For this reason, rectangular
duct should not be used in certain applications (e.g., with combustible dusts, etc.).
The equivalent diameter for non-circular ducts is calculated
using the following equation. The wetted perimeter noted in
the equation below is the inside perimeter of the odd-shaped
duct corresponding to its cross-sectional area.
dequiv = 1.3 (W H H)0.625 / (W + H)0.25
[9.5]
where: dequiv = equivalent diameter, in [mm]
W = duct width, in [mm]
H = duct height, in [mm]
This equation is also noted in Table 9-6 (IP and SI).
9.3.5 Friction Loss in Flexible Straight Duct. The duct
friction loss factor per foot [meter] for flexible duct with covered wires is shown to average:
F′d(flex-duct) = 0.0311(V0.604/Q0.639)
[9.6] IP
where: V = velocity, fpm
Q = airflow, acfm
In SI units, the equation is:
F′d(flex-duct) = 0.0186(V0.604/Q0.639)
[9.7] SI
where: V = velocity, m/s
Q = airflow, am3/s
Note that use of flexible duct segments should be avoided
except where necessary to connect hoods or process equipment to straight duct segments. This is because straight sections of flexible duct have almost twice the losses of similarly
sized metal duct. Note that the flexible duct friction loss factor
information is stated for straight duct lengths. Flexible duct, by
9-6
Industrial Ventilation
Local Exhaust Ventilation System Design Calculation Procedures
9-7
FIGURE 9-2. System duct calculation parameter location
its nature, is seldom straight. Typically, bends in flexible duct
can produce extremely large losses that cannot be easily predicted. If used, maintain flexible duct segments as straight and
as short as possible.
Additionally, Equations 9.6 and 9.7 do not reflect the wide
varieties of materials, wires and construction methods from
manufacturer to manufacturer. If flex duct is used in an LEV
system design, its manufacturer should be contacted to determine its actual duct friction loss factor per foot [meter] value.
9.3.6 Dynamic Losses in Elbows. The dynamic loss solely
due to turbulence generated from the redirection of airflow is
a product of the number of 90-degree elbows in the duct segment, an elbow loss factor(s) (Fel) and the duct velocity pressure, VPd.
Elbow Loss = (# of 90° elbows)(Fel)(VPd)
[9.8]
This loss is in units of inches-water gauge ("wg) [Pa].
Elbow loss factors are a function of the number of pieces
(gores) used to make the elbow and the radius-to-duct diameter (r/d) ratio. Stamped elbows and elbows made with more
gores, and those with higher r/d rations have resulted in lower
dynamic losses.
For example, the elbow loss factor, Fel, for a 5-piece 90degree elbow of radius/diameter (r/d) = 1.5 is shown to be 0.24
in Figure 9-e. A stamped elbow with r/d = 2.0 has an Fel of
0.13. When multiplied by the number of elbows and the duct
velocity pressure in that segment, the resulting value is the
dynamic elbow loss in "wg [Pa] (i.e., it represents the loss
associated solely with the change in direction; it does not
9-8
Industrial Ventilation
account for friction losses in the elbow).
9.3.7 Dynamic Losses in Branch Entries. Most LEV systems consist of a number of hoods connected by duct branch
segments to a main duct at a series of branch entries or junctions. The dynamic loss associated with joining each branch
duct’s airstream with the main duct’s airstream is the product
of a branch entry loss factor, Fen, multiplied by the duct velocity pressure, VPd. The branch entry loss factor is a function of
the branch entry’s geometry.
Branch Entry Loss = (Fen)(VPd)
[9.9]
This loss is notated in units of inches-water gauge ("wg) [Pa]
and is accounted for in branch only. See Figure 9-f for branch
entry loss factors.
9.4
STATIC PRESSURE LOSSES – SPECIAL
CONSIDERATIONS
9.4.1 Contractions and Expansions. Contractions are
used when the size of the duct must be reduced to route it
through a confined area, fit it to a piece of equipment or hood,
or to provide a high discharge velocity at the end of the stack.
Duct contractions result in an increase in the velocity of the
airstream and, therefore, an increase in its velocity pressure.
This will result in increased system resistance. Contractions
may also increase corrosion of duct segments and fittings
and/or generate excessive noise levels.
Expansions are used to fit a particular piece of equipment or
hood to a duct or to reduce the energy consumed in the system
by reducing the duct’s velocity and friction. Expansions are
not usually desirable in particulate systems since the duct
velocity may fall below the minimum transport velocity and
material may settle in the ducts.
The regain of pressure in an expansion, or loss of pressure
in a contraction, is possible because static pressure and velocity pressure are mutually convertible. This conversion is
accompanied by some energy loss or regain. The amount of
this loss or regain is a function of the geometry of the transition
piece (i.e., the more abrupt the change in velocity, the greater
the loss), and depends on whether air is accelerated or decelerated through the fitting.
Figure 9-3 displays plots of the changes in total and static
pressure through contractions and expansions. Reference Figure
9-d to determine how to calculate the pressure regain associated
with an expansion or loss associated with a contraction.
9.4.2 Special Expansion Considerations – Evasé
Discharge. An evasé discharge is a gradual enlargement at the
outlet of an LEV system (Figure 9-d). The purpose of the
evasé is to reduce efficiently the discharge air velocity; thus,
the available velocity pressure can be regained and credited to
the local exhaust system instead of being wasted. Practical
considerations usually limit the construction of an evasé to
approximately a 10° total angle (5° side angle) and a discharge
velocity of about 2,000 fpm (10 m/s) or a velocity pressure of
0.25 "wg (63 Pa) for normal LEV systems. Further streamlining or lengthening of the evasé yields diminishing returns.
However, for optimal vertical dispersion of contaminated
air, many designers consider that discharge velocity from the
stack should not be less than 3,000 fpm (15 m/s) to 3,500 fpm
(18 m/s). When these considerations prevail, the use of an
evasé is questionable. Additionally, the structural requirements
for the support of an evasé may add more initial costs than can
be realized in energy savings over the life of the project.
Note that it is not necessary to locate the evasé directly after
the outlet of the fan. Further, depending on the evasé location,
the static pressure at the fan discharge may be below atmospheric (i.e., negative). Remembering that air flows from an
area of higher to lower total pressure, this is possible as long
as the duct velocity pressure, VPd, is greater than the static
pressure, SP, thereby yielding a positive total pressure, TP.
9.4.3 Determining the Loss in Traps and Settling
Chambers. Traps and settling chambers are used in LEV sys-
tems to remove sparks, embers or other particulate material
from the airstream. If properly designed, they are an effective
means of capturing larger particles. Figure 9-f indicates how to
determine the static pressure loss as air moves through a trap
or settling chamber. Note that settling chambers are not recommended for the collection of combustible dusts.
9.5
BASIC SYSTEM DESIGN PROCEDURES AND
CALCULATIONS
A simple LEV system is comprised of a hood, duct segment
and special fittings leading to and from an exhaust fan. A complex LEV system is merely an arrangement of several simple
local exhaust systems connected to a common duct called a
main. Airflow in the system may be measured or specified as
being at either actual (acfm) [am3/s] or standard conditions
(dscfm [nm3/s]). The following procedure is a basis for performing system design calculations:
1) Flow rate (Q) will be specified in acfm [am3/s] for all
base value calculations. If a flow rate is specified in
dscfm [nm3/s] it must be converted to acfm [am3/s]
using the appropriate density factor prior to proceeding
with the LEV system design.
2) Actual flow rate (acfm [am3/s]) will be used for the determination of the duct size (using appropriate minimum
duct transport velocities, as discussed in Chapter 5).
3) Actual flow rate (acfm [am3/s]) will be used to determine the velocity pressure in any LEV system segment.
4) Actual flow rate (acfm [am3/s]) will be used for the
specification and sizing of all air cleaning devices (see
Chapter 8) and fans (see Chapter 7).
5) Density factor (df) will be calculated for all appropriate
conditions.
6) Flow rate at standard conditions (dscfm [nm3/s]) will
be calculated only after the base flow rate in acfm
Local Exhaust Ventilation System Design Calculation Procedures
9-9
9-10
Industrial Ventilation
Local Exhaust Ventilation System Design Calculation Procedures
[am3/s] and density factor have been determined. Such
flow rates are necessary for some LEV system design
calculations involving heat, moisture and the combining of airstreams at different densities and moisture
contents. In many cases in LEV system design, it is not
necessary to calculate dscfm [nm3/s].
7) Absolute pressure will only be considered at the fan
when calculating the density factor for absolute pressure (dfp).
EXAMPLE PROBLEM 9-1 (Effects of Evasé) (IP Units)
Determine the effects of adding a 40"-long evasé to the
discharge of a centrifugal fan with the following conditions:
Point
d
Q
V
VP
SP
1 Fan Inlet
20
8,300
3,800
0.90
-7.27
8,300
3,715
0.86
2 Fan Discharge
(16.5" H 19.5")
3 Round Duct
Connection
(fan outlet)
20
3,800
0.90
4 Evasé Outlet
28
1,940
0.23
0
To calculate the effect of the evasé, see Figure 9-d for an
expansion at the end of the duct where the diameter ratio,
d4/d3 = 28/20 = 1.4 and taper length L/d3 = 40/20 = 2.0.
R = 0.52 H 70% (since the evasé is within
5 diameters of the fan outlet)
VP3 = 0.9 "wg
SPexp = -R(VP3) = -(0.52)(0.7)(0.9)
= -0.33 "wg
SP4 = 0" (i.e., atmospheric pressure at end of duct)
SP3 = SP4 + SPexp
= (0.0 "wg) + (-0.33 "wg)
= -0.33 "wg
FSP = (SPoutlet – SPinlet) – VPinlet
= -0.33 "wg – (-7.27 "wg) – 0.9 "wg = 6.04 "wg
If a ‘no-loss’ stack was added to the fan (see Chapter 5,
Figure 5-4) and the effects of the evasé were not considered,
then the fan static pressure, FSP (see Chapter 7 for
discussion of FSP), would have been:
FSP = (SPoutlet – SPinlet) – VPinlet
= -0.0 "wg – (-7.27 "wg) – 0.9 "wg = 6.37 "wg
or 5% higher than the fan with the evasé (resulting in a 5%
higher operating horsepower over the life of the installation).
Note that this result does not account for the friction loss for
the straight duct segment connecting the no-loss stack to the
fan outlet that will likely have negligible impact on the
operating horsepower.
9-11
9.5.1 Hood Airflow at Nonstandard (Actual) Conditions.
The successful use of an LEV system to control airborne contaminants requires the proper determination of hood
airflow(s). Appropriate airflows develop the velocities necessary to capture and carry contaminants into and through the
hood and then into duct systems.
In an LEV system moving air at nonstandard conditions, its
air is less dense than air at standard conditions. For example,
at elevations above sea level, the air is at a lower density. The
methods defined in Chapter 3 use the airstream’s actual conditions to determine its density, density factor and actual flow
rate (acfm) [am3/s].
When selecting a capture velocity based on the guidelines in
Chapter 6 (Table 6-2) to derive a hood airflow rate, the designer
should consider the upper end of the range when working with
large dust particles at high temperatures or elevation (> 100 F
[> 38 C] and/or > 5,000 feet [> 1500 m] above sea level).
The actual airflow rate (acfm) [am3/s] is necessary for sizing
of ducts, determining air/cloth ratio for fabric filters and providing the correct size of fan. A system’s mass flow rate
(pounds/per minute or dscfm) is required to determine air conditions (i.e., amount of moisture, enthalpy, etc.) for an
airstream. There are cases where either or both airflow rate
values (acfm and scfm [am3/s and nm3/s]) may be required.
Knowing the density factor (as a function of elevation, moisture, temperature, absolute pressure), as well as the moisture
content and heat load, will allow the calculation of actual conditions from standard conditions or vice versa (see Chapter 3,
Sections 3.7 and 3.9).
Airflow in acfm [am3/s] is calculated using:
Qact = (Qstd)(1 + ω)/(df)
[9.10]
where:
Qact = actual flow, acfm [am3/s]
Qstd = flow, dscfm [nm3/s]
ω = moisture content, (lbm H2O)/(lbm da)
[kg H2O/kg da]
df = density factor
Equation 9.10 can be rearranged to solve for the airflow rate
at standard conditions:
Qstd = [(Qact)(df)] /(1 + ω)
[9.11]
9-12
Industrial Ventilation
EXAMPLE PROBLEM 9-2 (ACFM Into Hood)
EXAMPLE PROBLEM 9-3 (DSCFM Calculation)
A hood designed as shown in VS-55-01 encloses a melting
furnace. The hood has a required capture velocity at all openings of 200 fpm [1.0 m/s] per the VS plate and the opening
sizes total 52 ft2 [4.83 m2]. The hood is located in a plant that is
4,300 feet [1,300 m] Above Sea Level (ASL) and the plant air
temperature going into the hood is assumed to be 70 F [21 C]
with no moisture. Calculate the required hood control airflow
rate in acfm from the VS plate requirements.
For the system in Example Problem 9-2, determine the airflow into the hood in standard conditions (dscfm). The density
factor (df) for the air in the plant at 70 F [21 C] and 4,300 ft
[1300 m] ASL is 0.86 (see Table 9-7, with interpolation or
Chapter 3, Equation 3.12 IP).
In this example, there is no moisture and the only effect on
the airstream density is the elevation since temperature is 70 F
[21 C]. As such, Equation 9.11 can now be solved:
Qstd = (10,400 acfm)(0.86)/(1 + 0.0) = 8,944 dscfm
Actual Airflow Rate = Q = AV = (52 ft2)(200 ft/min)
[Qstd = (4.83 am3/s)(0.86)/(1 + 0.0) = 4.15 nm3/s]
= 10,400 acfm
In metric units:
[Q = AV = (4.83 m2)(1.0 m/s) = 4.83 am3/s]
EXAMPLE PROBLEM 9-4 (Calculating Mass Flow Rate)
The procedure requires the change of the acfm back to scfm
for the beginning of the system design calculation procedure.
This allows for a base value to be manipulated by all density
conditions before designing the duct and other equipment.
After the airflow is selected from the hood requirements (using
information from VS plates in Chapter 13, Chapter 6 or other
process requirements), the value in acfm [am3/s] must be
returned to its standard conditions.
In the field, there are cases where determining an
airstream’s mass flow rate of dry air (ṁda) is required, particularly in processes involving moisture. The transfer of volumetric flow rate (acfm) [am3/s] to mass flow rate (pounds-mass of
dry air per minute; lbm da/min) [kg da/s] requires an understanding of the concept of density (see Chapter 3). To determine an airstream’s mass flow rate, it must first be converted
to standard conditions, as was done in Example Problem 9-2.
Mass flow rate (for dry air at standard conditions) is determined using the following equation:
ṁda = (rstd)(Qstd)
Determine the mass flow rate of the airstream (ṁa) in
Example Problem 9-3 (pounds-mass of dry air per minute).
ṁa = (0.075 lbm da/ft3)(8,944 ft3/min) = 670.8 lbm da/min
[ṁa = (1.204 kg da/m3)(4.15 nm3/s) = 5.00 kg da/s]
In Example Problems 9-2 through 9-4, it was determined
that the actual airflow rate into the melting furnace hood is
10,400 acfm [4.83 am3/s] (from VS plate), that calculates to
8,944 scfm [4.15 nm3/s] and 670.8 lbm da/min [5.00 kg da/s].
9.5.2 Addition of Materials Inside the Hood. In some
cases, an enclosed process may add gases or moisture to the
calculated control airflow going into the face of the hood.
These materials must be accounted for in the calculation of the
connected duct system in order to properly size the duct and
air handling equipment.
[9.12]
where:
ṁda = mass flow, lbm/min [kg/s]
Pstd = standard density, lbm/ft3 [kg/m3]
Qstd = flow, dscfm [nm3/s]
EXAMPLE PROBLEM 9-5 (Density Change Inside Hood)
The melting furnace in Example Problem 9-2 is generating
3,000 acfm [1.42 am3/s] of gases at 1,900 F [1038 C] with no
moisture, that must also be controlled by its exhaust hood. The
standard density of this process gas is the same as air (0.075
lbm/ft3) [1.204 kg/m3]. Determine the total pounds of material
(air plus gases) exiting at the melting furnace hood’s duct connection.
It was determined in Example Problem 9-4 that the air coming into the hood from the plant (ṁa) totals 670.8 lbm da/min
[5.00 kg da/s]. The gases being generated inside the hood (ṁf)
Local Exhaust Ventilation System Design Calculation Procedures
must be added to this value. The density factor for the gas at
1,900 F [1038 C] (see Chapter 3, Equation 3.14 IP) is:
dft = (ract)/(rstd) = (Tstd)/(Tact) = (70 + 460)/
(1,900 + 460) = 0.22
EXAMPLE PROBLEM 9-6 (Combining of Airstreams)
(IP Units)
Determine the exit temperature, density factor and airflow
rate of the combination of hot and cold gases coming from the
enclosure defined in Example Problems 9-2 through 9-5.
[dft = (ract)/(rstd) = (Tstd)/(Tact) = (21 + 273)/
(1038 + 273) = 0.22]
This value is determined by the process requirements and,
therefore, is independent of the elevation of the plant where the
furnace is located. Therefore, dft would be the only consideration in determining the volume of hot gas generated by the
melting furnace. Solving Equation 9.11 for the standard conditions:
The mass flow rate of 70 F [21 C] air coming through the
face of the hood (ṁa) was determined to be 670.8 lbm da/min
[5.00 kg da/s]. The mass flow rate of the furnace exhaust gases
(ṁf) was determined to be 49.5 lbm da/min [0.37 kg da/s] at
1,900 F [1038 C].
Rearranging Equation 9.16 to solve for the temperature of
the combination of gases (Tcomb):
Qstd = (Qact)(df)/(1 + ω) = (3,000 acfm)(0.22)/(1 + 0.0)
= 660 scfm
Tcomb =
[Qstd = (Qact)(df)/(1 + ω) = (1.42 am3/s)(0.22)/(1 + 0.0)
= 0.31 nm3/s]
=
{(670.8)(70) + (49.5)(1,900)}/(720.3)
=
196 F
From Equation 9.12:
Determining the total mass flow rate (ṁtotal) of the extra
gases generated by the process itself and those entering
through the hood face is now a simple addition of masses:
91 C]
Rearranging Equation 9.12 and solving for the standard airflow rate of the combined gas stream, Qstd:
Qstd = (ṁcomb)/(rstd) = (720.3)/(0.075) = 9,604 scfm
NOTE: This is the sum of scfm [nm3/s] from both airstreams.
[ṁa + ṁf = ṁtotal = (5.00 + 0.37) kg/s = 5.37 kg/s]
9.5.3 Combining of Gases of Different Densities Due to
Temperature. Example Problems 9-2 through 9-5 in Section
9.4 show the effects of density and how to combine the mass
flow rates of two gas streams. The principle for this combination is the Law of the Conservation of Mass (see Chapter 3,
Section 3.7).
[9.13]
Assuming that there is no heat loss through the walls of the
hood, there is also a Conservation of Energy in this system
(see Chapter 3, Section 3.8).
[9.14]
For an ideal gas that contains no moisture (see Chapter 3,
Section 3.4), Equation 9.14 is rewritten:
[9.15]
Cancelling specific heat, CP, from the equation yields:
(ṁa)(Ta) + (ṁb)(Tb) = (ṁc)(Tc)
[(5.0)(21) + (0.37)(1038)]/(5.37)
[Qnormal = (ṁcomb)/(rstd) = (5.37)/(1.204) = 4.46 nm3/s]
ṁa + ṁf = ṁtotal = (670.8 + 49.5) lbm/min
= 720.3 lbm/min
(ṁa)(CP)(Ta) + (ṁb)(CP)(Tb) = (ṁc)(CP)(Tc)
{(ṁa)(Ta) + (ṁb)(Tb)}/(ṁcomb)
The conditions of the combination leaving the hood: 720.3
lbm/min @ 196 F [5.37 kg/s @ 91 C]
[ṁf = (1.204 kg/m3)(0.31 nm3/s) = 0.37 kg/s]
ṁa(ha) + ṁb(hb) = ṁc(hc)
[Tcomb =
=
ṁf = (0.075 lbm/ft3)(660 dscfm) = 49.5 lbm/min
ṁa + ṁb = ṁc
9-13
[9.16]
NOTE: All temperatures must be in consistent units of
measure.
Information for calculating df is shown in Chapter 3, Section
3.5. There are two items affecting density of the gases exiting
the furnace in these examples: 1) the gas is at an elevated temperature 196 F [91 C], and 2) the hood is located at 4,000' ASL
[1200 m]. The density factor for elevation was determined previously (0.86) in Example Problem 9-3. The density factor for
temperature is calculated using Chapter 3, Equation 3.14:
dft = (Tstd)/(Tact) = (460 + 70)/(460 + 196) = 0.81
[dft = (Tstd)/(Tact) = (273 + 21)/(273 + 91) = 0.81]
The density factor of the combination considering both temperature and elevation is:
df = (dft)(dfe) = (0.81)(0.86) = 0.70
The actual airflow rate is determined by Equation 9.10:
Qact = (Qstd)(1 + ω)/(df) = (9,604)(1 + 0.0)/(0.70)
= 13,720 acfm @ 196 F @ 4,000' ASL
[Qact = (4.46)(1 + 0.0)/(0.70) = 6.37 am3/s @ 91 C
@ 1200 m ASL]
NOTE: Whenever airflow is specified in actual conditions it
is important to list the conditions immediately following (196 F
and 4,000' ASL) [91 C and 1200 m ASL]. This is not required
when listing standard [or normal] conditions although it is good
practice to provide a notation of the conditions whenever defining flow.
9-14
9.6
Industrial Ventilation
CALCULATION SHEET DESIGN PROCEDURE
System design usually considers only the conditions at initial start-up and installation (i.e., it defines a single point of
operation). However, the system itself is dynamic and continuously changing. This results in fluctuating readings for volume and pressure at any point in the system over time.
Readings taken at start-up and commissioning may not be
repeated again as the system ages (see Appendix C, Testing
and Measurement of Ventilation Systems). After the system is
in use, it will lose some effectiveness as dust covers the duct
walls (changing friction losses) and fan impellers and equipment begins to wear. As such, the designer must consider the
conditions during the operating life of the system.
The calculation procedure to determine the SSP for an LEV
system is a continuing/iterative process and does not end with
the first system design solution. The design process might be
repeated several times – from the original conceptual design to
the final drive speed specification from as-built drawings.
An LEV system designer must have a thorough understanding of all design principles prior to attempting to use a calculation sheet (calc sheet). The calc sheet is a useful tool for identifying the values for various inputs and calculations necessary
to derive the duct sizes, SSP and fan requirements for such systems. It may also be used to identify ducts with very high
velocities that could wear prematurely, and to analyze the
branches with the highest pressure drop to identify where system pressure could potentially be reduced. An air balance technician may also use data from it during system commissioning
and balancing.
While a tremendous aid during the system design process,
the calc sheet should not be relied on as a means of predicting
the exact operating conditions in all branches throughout the
life of an LEV system. Additionally, one must not simply consider the calc sheet a tool only to be used for sizing ducts or
selecting an air cleaning device or a fan.
9.6.1 Use of the Velocity Pressure Method. This procedure uses the Velocity Pressure Method to determine duct sizes
and fan conditions (see Sections 9.3 and 9.6). Typically, one
begins the design process either at the hood and corresponding
duct segment located farthest from the fan and proceeds systematically from there. Pertinent airstream characteristics for
the segment are determined and entered, the duct is sized (if
appropriate), and the velocity pressure is calculated.
The designer then determines the applicable hood entry and
system component loss factors (i.e., Fh, Fd, Fel, Fen, etc.) and
enters them on the calc sheet. If appropriate, the hood static
pressure is calculated. Then, loss factors are totaled and the
sum is multiplied by the velocity pressure in that segment to
obtain the actual losses in "wg [Pa]. Other appropriate losses
are also determined and all segment losses are summed to
yield a segment pressure loss. The system’s various segment
static pressure losses are then summed, as appropriate, until
one can determine an SSP.
The calculations necessary to derive the SSP may be documented long-hand, or more concisely in a calc sheet (Figure 9-6).
9.6.2 Use of the Calculation Sheet. The calc sheet is built
as a series of columns (normally one column for each duct segment) and rows (data for a particular column). The cell location for entered value is made using a matrix notation. The first
value in a cell identifier would be the column (e.g., A, B, etc.)
and the second value would be the row (e.g., 1, 2, etc.). For
example, in Sample System Design #1 (Figure 9-6 (IP)), the
value at cell A/24 would be 2.1 "wg. It is found at the intersection of Column A and Row 24. Similarly, the value in cell D/11
would be 5" diameter.
When using the calc sheet, work from the top to the bottom
of each column, compiling data for a given system segment
that is identified by its segment identification. The designer
inputs known data from sketches, VS plates and other
resources into the appropriate rows at the top of the calc sheet.
Note that certain rows (1, 2, 3, 4, 5, 6, 11, 12, etc.) contain
asterisks next to the row number. This asterisk indicates data
entry points necessary for the design in certain cases. Other
row values are normally calculated from these input values.
The calc sheet also includes shaded rows (5, 6, 7, 8, and 9).
These are usually required only when non-standard air is
encountered (i.e., when df does not equal 1.0 and/or a value for
total heat is noted).
Data entry points can all be inserted into the calc sheet
before doing calculations for the column or can be entered as
the calculations proceed down a column. In either case, a
series of calculations are performed working down from the
top of the column to obtain a segment static pressure loss
(resistance) for that segment (Row 38).
Once a segment (column) is complete, the designer then
moves to the next segment (column) and the process begins
again. If that segment is a branch duct meeting at a junction
with the main duct, the magnitude of the two segment static
pressure losses (from Row 38) are compared. If the pressures
are not equal, the pressure that is greater in magnitude (e.g., -4
is greater in magnitude than -2 and 4 is greater than 2) is
recorded as the governing pressure in Row 39 of the segment
with the pressure that is lower in magnitude (see Cell A/39 in
Figure 9-6 (IP)). Adjustments are then made as required to balance the two branches so that there is only one static pressure
at the junction (see Section 9.9). Once balanced and appropriate system modifications have been noted, the airflows are
added and the design process proceeds to the next segment.
When no branch entry is encountered but additional duct
segments occur, consecutive segment static pressure losses are
added to yield a cumulative static pressure (Row 40).
The calculation continues until the fan segment is reached
where the inlet static pressure (SPin) to the fan is determined.
The same procedure starts beginning with the outlet of the fan
and an outlet static pressure to the fan (SPout)is determined. After
Local Exhaust Ventilation System Design Calculation Procedures
the inlet velocity pressure and inlet and outlet static pressures are
determined for the system, a system static pressure (SSP) can be
calculated and fan static pressure (FSP) can be specified.
9.6.3 Calculation and Input of System Design Data on
the Calculation Sheet. First, note the elevation for the plant
location and input the value in the title block area of the calc
sheet (“z”). This will be used to calculate the density factor for
elevation (dfe) in the plant (see Chapter 3, Section 3.5). Then
input all other pertinent data for the system in the appropriate
places in the title block.
The designer should then start with the hood and corresponding duct segment that travels the greatest distance to the fan.
The design may also begin at the hood/duct segment that will
yield the greatest pressure loss. A segment is defined as the
constant diameter round (or constant area rectangular) duct
that separates points of interest such as hoods, branch entry
points, fan inlet, etc. A segment may also identify the entry and
exit points to an air cleaning device or a fan. Once the initial
system segment, or “main”, has been determined:
1) Select a segment identification; this is usually specified
by a number (1, 2, etc.) for the hood combined with a
single letter for the junction at the end of the segment
(A, B, etc.). The first identification input is entered into
cell A/1.
2) If the column involves a hood design or other source of
air (bleed-in, etc.), select an airflow rate based on the
toxicity, physical and chemical characteristics of the
material and the ergonomics of the process. Values for
airflow rates, minimum transport velocities and hood
entry loss factors are located in Chapters 5, 6, 9 and 13.
Actual duct airflow rates (Qact) are input in Row 3.
3) Maintaining the minimum transport velocity is critical
for systems transporting particulate, condensing vapors
or mist and to prevent explosive concentrations from
building up in the duct. Chapter 5, Section 5.2.1, provides information on ranges of duct velocities for the
transporting gases and vapors. Input the value for the
minimum transport velocity into Row 4 (see Chapter 5,
Table 5-1). Hood entry to duct loss factors determined
from VS plates in Chapter 13 can be immediately input
into Row 16 (if a compound or orifice style hood)
and/or Row 20.
4) Account for the volume of contaminants generated
inside the hood enclosure – defined in scfm @ 0.075
lbm/ft3 [1.204 kg/m3], any moisture added and the total
heat of the airstream. Note that this may differ from the
actual contaminants being generated and the designer
will be required to re-state these contaminants in terms
of scfm of air (Row 8). The calc sheet uses acfm as a
start point (Row 3) because the face velocities and airflow going into the hood are at local conditions (acfm)
[am3/s]. This allows for the determination of the appropriate density factor for use in all of the eventual calcu-
9-15
lations for that branch (see Chapter 3, Section 3.5).
5) Calculate the branch density factor (Row 7) considering the effects of elevation, temperature, moisture and
(if appropriate) absolute pressure for the airstream
coming from the hood (see Chapter 3). The density factor must be used for proper sizing of the duct. NOTE:
The density factor is affected by the absolute pressure
inside the duct. However, for most calculations, the
absolute pressure will only be considered at the fan
inlet. This is where the effects are usually the greatest
and the information is needed to specify the fan. If a
more detailed system calculation is to be considered or
there are very high or low pressures throughout the system (« ± 20 "wg) [« ± 5 kPa], then the designer may opt
to consider these effects in additional system segments.
6) Calculate scfm [nm3/s], if required, for combining of
airstreams with two different temperatures (Row 8).
7) Determine the target duct area by dividing the actual
duct flow rate (Row 3) by the minimum transport
velocity (Row 4) to determine a target duct area (Row
10). Then convert the target duct area into the selected
duct diameter (Row 11). A commercially available duct
size (Table 9-2) should be selected.
Remember, if solid or liquid particulate or condensable
vapor is being transported through the system, a minimum transport velocity must be maintained (see
Chapter 13 and Chapter 5, Table 5-1). For such systems, if the target duct area does not match the diameter
of a standard duct size, select the next smaller size (from
Table 9-2 to ensure that the actual duct velocity is equal
to or greater than the minimum required.
8) Calculate the duct velocity (Row 13) and corresponding velocity pressure (VP) (Row 14). Remember to
include the effects of density in calculating the VP.
9) Determine the absolute value of the hood static pressure and input its value into Row 24. Use of the
absolute value aids in properly determining the segment pressure loss (Row 38).
10) Using the line sketch, determine the straight duct
length (Row 25) for the duct segment and the number
and type of elbows needed. The straight duct length is
the centerline distance along the duct (the distance
between the intersections of the centerlines of the duct
components in the segment). Input values respectively
into Rows 25 and 27.
11) Determine the duct friction loss factor per foot [meter]
(F′d) of duct either by calculation, or use of Table 9-5
(IP & SI), or use of Figures 9-b and 9-c and input into
Row 26. Additionally, determine the type of elbows
and, if applicable, branch entry included in the segment
and input values for their respective loss factors (see
Figures 9-e and 9-f) into Rows 28 and 29. When nec-
9-16
Industrial Ventilation
essary, enter the special fitting loss factor into Row 30.
12) Calculate the static pressure losses for the duct segments that merge at a common junction point.
13) Calculate the condition of the air at each branch by
considering moisture, heat and mass flow in the combination from the two branches and balancing mass
(dscfm) [nm3/s] or lbm da/min [kg da/s], moisture and
heat. Review these conditions to ensure that the air is
safely above the dew point if moisture is present from
the process. Use the combined air conditions for
designing the next segment.
14) Directly at each junction point, there will be one and
only one value for static pressure (SP), regardless of
the path taken to reach that point. If not ensured by the
design process, the system will self-balance by reducing the flow rate in the higher-resistance duct segment(s) and increasing the flow rate in the lower-resistance duct segment(s) until there is a single SP in the
duct downstream of each junction point. Balancing the
SP at any junction point can be achieved in the design
process by use of the balance-by-design or blast
gate/orifice plate method. See Section 9.8 and Chapter
4, Section 4.5 for a further discussion of these methods
for balancing static pressure at a junction.
Select both the air cleaning device and fan based on the final
calculated system airflow rate in acfm [am3/s] (considering
temperature, elevation, static pressure, moisture condition,
contaminant and heat loading, physical and chemical characteristics, and overall system resistance).
Check the duct sizes designed against the available space
and resolve any interference problems (i.e., will the elbow or
duct size desired actually fit into the available space). This
may cause a redesign of a part of the system. Consider fan inlet
and outlet conditions and the system effects that will derate the
fan (see Chapter 7, Section 7.4).
EXAMPLE PROBLEM 9-7 (Input to Calculation Sheet)
(IP Units)
Input the data for the hood in Example Problems 9-2 through
9-6 into an ACGIH® Calculation Sheet (Figure 9-4). Note the
method of entering data from top to bottom. First, the elevation
of the system (4,300' ASL) is added to title block area of the
sheet (“z”). This is the reference for the calculation of density
factor due to elevation (dfe). Then a segment identification
number is assigned by the designer (Row 1). This usually
includes a start and end value separated by a hyphen. In this
case, 1-A indicates a hood (designated by a number) and the
“A” is the end point of the duct connected to the next duct segment. This is placed in cell A/1. Other numbers, letters or com-
binations may be used for hoods, including machine names or
numbers.
The remaining data are entered vertically down the column
into individual cells. These cells are identified by a matrix designation (see Section 9.6.2). The dry bulb temperature of the
combination from the hood was calculated in Example Problem
9-6 and its value (196 F) inserted into cell A/2 (Column A; Row
2). Similarly, the values for dscfm (9,604; from Example
Problem 9-6), pounds of dry air per minute (720.3 from
Example Problem 9-5), df (0.70 from Example Problem 9-6)
and actual duct flow rate (13,720 acfm; from Example Problem
9-6) were added to their respective cells.
Note that simple systems (no heat or moisture and elevation
below 1,000 ft ASL) may have values simply transferred from
the VS plates (Chapter 13) or other calculated values and that
more complicated values of acfm (involving heat, moisture and
elevation, etc.) may require calculation on a separate sheet. As
mentioned previously, these simple systems do not require that
information be placed into the shaded areas of the sheet. The
equations referenced on the calc sheet are shown on the right
edge of the calc sheet (Figure 9-6).
9.7
SAMPLE SYSTEM DESIGN #1 (SINGLE-BRANCH
SYSTEM AT STANDARD AIR CONDITIONS)
NOTE: The solutions to this problem in imperial (IP) and
metric (SI) units are each unique and do not correspond in size
or description.
Figures 9-5 (IP & SI) show a simple ventilation system with
a single hood. These figures provide graphical representations
through the system showing the magnitude and relationships
of total, static and velocity pressures on both the inlet and the
outlet sides of the fan. It should be noted that velocity pressure
(VP) is always positive. Total and static pressure may be either
negative or positive with respect to atmospheric pressure.
Total pressure (TP) is always greater than static pressure (SP)
(i.e., TP = SP + VP). Also note that VP can be affected by the
air conditions (moisture, temperature, elevation and pressure),
but in this example standard air (df = 1.0 and ω = 0.0) is considered.
The following steps refer to the ACGIH® Calculation Sheets
(calc sheet) shown in Figures 9-6 (IP & SI). Data are entered in
rows denoted with an asterisk. The other rows require calculations to determine their inputs. Not all rows need to be used
based on the requirements of the system (i.e., if there are no
elbows in the system then no data are required in Rows 27, 28,
and 32).
Step 1. In the column for the first duct segment (from the
hood at “1” to the inlet of the filter at “A”), name the duct segment (1-A) and place in cell A/1. Since the air in this problem
Local Exhaust Ventilation System Design Calculation Procedures
9-17
FIGURE 9-4. Data entry to calculation sheet (Example Problem 9-7)
is at standard conditions, input 70 F [21 C] for the air temperature in cell A/2.
Step 2. Input the required airflow (in acfm) [am3/s] into cell
A/3. This value comes from the information in the VS plate
(VS-80-11) for a grinding wheel hood in Chapter 13. From the
same VS plate, input the minimum transport velocity (4,000
fpm) [20 m/s] into cell A/4.
Step 3. Determining the duct size is a two-stage process.
When 390 acfm [0.25 am3/s] is carried exactly at 4,000 fpm [20
m/s], the duct area required is 0.098 square feet [0.013 m2] (A =
Q/V) and is shown in Cell A/10. Solving for duct diameter at
that area yields a value of 4.23" [128 mm]. Since this size duct
is impractical for fabricators, a more standard size is considered.
If the designer opts for a larger duct, the velocity would not
meet the minimum requirement of 4,000 fpm [20 m/s]. Thus,
the next smaller commercially available duct is chosen – in
this case 4.0" [120 mm] diameter; that value is entered in Cell
A/11. The area for a 4" diameter duct equals 0.087 square feet
[0.011 m2] (Table 9-2) that yields a velocity of 4,483 fpm
[22.73 m/s]. These data are inserted in Rows A/12 and A/13,
respectively.
Additionally, remember that the SPh is the sum of the
dynamic hood losses and the energy transfer as air moves from
stillness outside the hood to the energy as it travels at the
velocity in the duct (Fa × VPd = 1VPd). The energy transfer is
commonly referred to as the acceleration loss and is discussed
in Section 3.6 of Chapter 3. Determine SPh from the equations
in Chapter 6, or available information in Chapter 13 (VS-8011 for this problem indicates that Fh = 0.65).
Given that this hood contains no slots (Rows 15–19), the Fh
is entered into Cell A/20. It is then multiplied by the duct velocity pressure (VPd) and the product is entered into Cell A/21.
Step 6. If there are other losses in the hood (e.g., a filter section or spray section that had resistance) they would be noted
in Cell A/23 and accounted for in the SPh. Since this system
does not have any other losses, SPh can be determined by
adding the values of Rows 20 and 22 (0.8 + 1.25) [202 + 310].
The absolute value of the SPh, 2.1 "wg [512 Pa], is input into
Cell A/24.
(The following steps add any other cumulative losses as the
system design proceeds to point “A”. These include the losses
due to straight duct, elbows, contractions, expansions, etc. In
this Example Problem there is only straight duct.)
Step 4. The velocity pressure in the duct is determined either
from Equation 3.17a, Equation 4 on the right side of the calc
sheet, or Table 9-4 (if standard air). This velocity pressure is
placed in Cell A/14. This completes all of the basic system
data entry for this segment and now the static pressure losses
can be calculated.
Step 7. From the drawing information in Figure 9-5, the
length of straight duct, 15' [4.5 m], is input into Cell A/25. The
duct friction loss factor per foot (F'd) is determined by use of
either Table 9-5 (IP or SI) or by using Equation 8 on the calc
sheet. The product of the duct length and F'd yields the duct
friction loss factor (Fd) which is inserted into Cell A/31.
Step 5. The first component of the system loss to address in
this segment is the hood static pressure (SPh). By definition
SPh is a negative value. However, in order to ensure the proper
summation of the losses for a given segment, it is recorded as
an absolute value on the calc sheet.
Step 8. Determine the number and type of fittings in the
duct segment. For each fitting type, determine its loss factor
(Figures 9-d, 9-e, and 9-f) and, when appropriate, multiply by
the number of fittings (as mentioned above, there were none in
this example). Input the data into Rows 27 through 30. All of
9-18
Industrial Ventilation
Local Exhaust Ventilation System Design Calculation Procedures
9-19
FIGURE 9-6 (IP). Velocity Pressure Method Calculation Sheet
9-20
Industrial Ventilation
FIGURE 9-6 (SI). Velocity Pressure Method Calculation Sheet
Local Exhaust Ventilation System Design Calculation Procedures
9-21
9-22
Industrial Ventilation
the factors for the components for the segment are compiled in
Cell A/33.
Step 9. Multiply the total in A/33 by the VPd (Cell A/14).
This is the total duct loss for all components in inches of water
[Pascals] for the duct segment and placed in A/34.
Step 10. Add the result of Steps 6 and 9. This combines the
hood and duct losses for the segment. If there are any additional losses (expressed in "wg [Pa]), such as for special fittings or
a velocity increase at a junction, include them also in the
appropriate location (Cells A/35 through 37). This establishes
the cumulative energy required, expressed as the segment static pressure loss, to move the design airflow through the duct
segment; it is input into Cell A/38. Note that the values recorded anywhere in Row 38 cells are recorded as negative (-) values if the segments occur prior to the fan in a single fan system; they are recorded as positive (+) values if the segments
occur after the fan in a single fan system. Given that it is the
first segment of the system, it is recommended that this value
also be entered into Cell A/40, the cumulative static pressure.
NOTE: The value in Cells A/38 through A/40 are most commonly negative when located prior to the fan inlet. This represents the value for the static pressure in the duct with respect
to that of air in the plant (0 "wg [0 Pa]). Similarly, values representing static pressure after the fan outlet are most commonly positive when compared to the plant air and would be
entered as such on the calc sheet. The value of -3.4" [-779 Pa],
in Cell A/40, would be used to begin the system calculation
and is close to representing a value that would be seen if a
measurement of pressure were taken at point “A”. (See
Appendix C for measurement methods.) The value represents
the static pressure required to pull 390 acfm [0.25 am3/s] @ df
= 1 through the duct and hood as designed.
The rest of the system design proceeds similarly with inputs
placed in columns adjacent to Column A. The second column
is designated as “A-B” and covers the next segment of the system, the fabric filter in this case. Because the only loss given
in the example is across the filter media (sometimes called
ΔP), 2 "wg [500 Pa], this information is placed in the cell designated for other losses (Cell B/35). This 2.0 "wg [500 Pa] is
again noted in Cell A/38 for the segment static pressure loss
(as a negative value because it also is located prior to the fan
inlet). The value for the segment static pressure loss from “A”
to “B” (Cell B/38) is added to the losses already accumulated
from “1” to “A” (Cell A/40) to arrive at -5.4 "wg [-1279 Pa]
and inserted into Cell B/40.
The design process continues for all segments up to the fan
inlet. Then the accumulation of negative static pressure is noted
in Row 40 (shown at Cell C/40) and is designated as static pressure into the fan inlet (SPi). Note that the losses in Column C
include another hood loss. This is because the air slowed to a
very low velocity in the collector and must be reaccelerated as
it enters the duct leading to the fan.
A new set of pressures are calculated on the positive or dis-
charge side of the fan. These are shown cumulatively in the same
row (in this problem in Cell D/40) and are designated as static
pressure at the fan outlet (SPo). The SPi and SPo, as well as the
velocity pressure going into the fan (VPi) (as shown in Cell
C/14) are used to calculate the system static pressure (SSP). This
is further explained in Section 9.10 and its calculation may be
recorded in the Notes section at the bottom of the calc sheet.
9.8
DISTRIBUTION OF AIRFLOW IN A MULTI-BRANCH
DUCT SYSTEM
Most LEV systems have multiple branches. In such systems, care must be taken to provide the correct balance of
flows at each hood and pressures at the junction of converging
branches. A junction can have only one static pressure at all
connected branches. Air will always take the path of least
resistance, a natural balance at each junction will occur if the
system is not balanced during the design phase. That is, the
exhaust flow rate will distribute itself automatically according
to the pressure losses in each branch. Therefore, it is necessary
to provide a means of distributing airflow between the branches in these systems in order to balance the static pressure at a
junction.
Balancing ensures that the required airflow at each hood
will never fall below the minimums specified in Chapters 6
and/or 13. Balancing the system also aids in the proper sizing
of duct systems and helps to prevent the settling of particulate
in the system.
There are two methods commonly used to achieve this balance:
a) Balance-by-Design Method: Also called the static pressure balance method. This method adjusts the flow rate
through the branch(es) until the static pressures at the
junction point are equalized. A reduction in the duct
size and/or change in fittings in the branch with the
lower resistance are often used to achieve the SP balance.
b) Blast Gate/Orifice Plate Method: This method increases
the static pressure in the lower resistance duct segment(s)
by means of some artificial device such as a blast gate,
orifice plate or other obstruction in the segment.
See Chapter 4, Section 4.5 for more information regarding
these two balancing methods.
One should also investigate whether the system static pressure (SSP) can be reduced by increasing the flow at one or
both hoods or by increasing duct sizes. When selecting the
most appropriate balancing method, the designer should consider the effect on total system horsepower and capital costs.
Consider a system consisting of one branch that includes
Hood A and its associated duct components. It has a static
pressure (SP) requirement of -3.5 "wg [-875 Pa]. An adjoining
branch containing Hood B and its duct components have an SP
requirement of -5.0 "wg [-1250 Pa]. When one compares the
Local Exhaust Ventilation System Design Calculation Procedures
magnitude of the two static pressures (e.g., -4 is greater in
magnitude than -2 and 4 is greater than 2), the fan must be able
to deliver a static pressure of -5.0 "wg [-1250 Pa] at the junction. Otherwise, Hood B will not have enough energy to generate all of its required design flow. This pressure at the junction with the greater magnitude is called the “governing pressure” and its branch is called the “governing branch.” If the fan
is selected to provide -5.0 "wg [-1250 Pa] there is now excess
pressure pulling at Hood A. The designer must provide 1.5
"wg (5.0" – 3.5") [1250 – 875 = 375 Pa] more resistance in the
branch serving Hood A to balance the flow conditions at the
junction.
Use of the balance-by-design method in this example creates additional resistance by decreasing the duct size of the
branch with the lower SP, thereby increasing frictional and
dynamic losses and raising the duct’s SP. Appropriately resizing the duct will generate the additional SP (in this case, 1.5"
[375 Pa]) to balance the two branches at their junction.
The blast gate/orifice plate method uses a partially closed
gate or an orifice plate to add the 1.5 "wg [375 Pa] of resistance.
When close to the proper setting, the gate should be able to meet
the requirements for balancing the branches. Remember to
secure blast gate settings so that they cannot be easily adjusted,
thereby defeating the balance achieved at the junction.
In either case, the object of both methods is the same: to
obtain the desired flow rate at each hood in the system while
maintaining the desired velocity in each branch and the main.
9.8.1 Use of the Balance-by-Design (Static Pressure
Balance) Method. In this type of design method, the calcula-
tion usually begins at the hood farthest from the fan in terms
of number of duct segments and proceeds, segment by segment, towards the fan. At each junction, the SP necessary to
achieve the desired airflow in one branch must equal the SP in
the joining branch (i.e., a comparison of the magnitude of the
SP in each branch is made). The SP in the branch whose resistance is greater in magnitude (i.e., the governing branch) is
referred to as the governing SP (SPgov). The SP in the branch
whose resistance is lower in magnitude is referred to as the
lower static pressure (SPlower) and its flow rate is designated as
the lower flow rate (Qlower).
When the ratio of SPgov to SPlower, or (SPgov/SPlower) is:
a) Greater than 1.2: redesign of the branch with the lower
SP should be considered. This may include a change of
duct size, selection of different fittings and/or modifications to the hood design.
b) Unequal but less than 1.2: balance can be obtained by
increasing the airflow through the branch with the
lower SP. This change in flow rate is calculated by noting that pressure losses vary with the square of the flow
rate (see Chapter 7, Section 7.5). Therefore:
Qcorr = Qlower(SPgov/SPlower)0.5
[9.17]
9-23
where:
Qcorr = corrected volumetric flow rate, acfm [am3/s]
Qlower = volumetric flow rate of branch whose
resistance is lower in magnitude (i.e., the
branch whose flow rate is to be modified),
acfm [am3/s]
SPgov = the static pressure of the branch whose
resistance is greatest in magnitude, "wg [Pa]
SPlower = the static pressure of the branch whose
resistance is lowest in magnitude, "wg [Pa]
NOTE: The square root of the SP comparison is always
greater than 1.0.
9.8.2 Use of the Blast Gate/Orifice Plate Method. This
design procedure depends on the use of blast gates and/or orifice plates located in branches or mains to provide additional
resistance to balance static pressures. Blast gates, also called
cut-offs, or slide gates (see Chapter 5, Figure 5-12) must be
adjusted after installation in order to achieve the desired flow
at each hood. At each junction the flow rates of two joining
ducts are achieved by adjusting a blast gate in one or more
branches to achieve the desired static pressure balance. It
should be noted that, once the system balance has been
achieved, a change in any blast gate setting in a system could
change the flow rates in all of the other branches. Readjusting
the blast gates can also result in increases to the actual FSP and
power requirements. Therefore, once the system balance has
been achieved, blast gates should be secured in place to prevent unauthorized alteration of system performance.
Similarly, orifice plates may be sized to reflect actual
requirements at start-up or when system revisions are made.
Their design usually infers a more permanent installation with
less chance of operator adjustment.
NOTE: The corrosiveness or abrasiveness of the airstream
should also be considered when using the blast gate/orifice
plate method.
Data and calculations involved for this method are the same
as for the balance-by-design method except that the duct sizes,
fittings and flow rates are not adjusted; the blast gates are set
after installation to provide the required design flow rates.
NOTE: Systems are commonly designed assuming that only
a fraction of the total number of hoods will be used at any given
time. In such instances, the flow to the branches not used could
be shut off with dampers. For LEV systems where particulate is
transported, this practice may lead to plugging in the main and
branch ducts due to settled particulate. This procedure is not
recommended unless the minimum transport velocity can be
assured in all ducts during any variation of opened and closed
blast gates (and may not be permitted if the particulate is considered a combustible dust). It is better to design these systems
with individual branch lines all converging very close to the fan
inlet so that lengths of duct mains are minimized or use a
plenum system design (see Chapter 4, Section 4.6.2).
9-24
9.9
Industrial Ventilation
INCREASING VELOCITY THROUGH A JUNCTION
(WEIGHTED AVERAGE VELOCITY PRESSURE)
Variations in duct velocity occur at many locations in LEV
systems. Small accelerations and decelerations are usually
compensated for automatically in the system where good
design practices and proper fittings are used.
Sometimes the final main duct velocity immediately following a junction exceeds the weighted average of the two
velocities in the branches entering the main. Air speed cannot
be increased through the fitting without an expenditure of
kinetic energy. If the difference between the weighted average
of the branch velocities and the velocity downstream following the junction is greater than zero, additional static pressure
is required to produce the increased velocity. This extra loss is
shown in Row 36 of the calc sheet. In previous editions, this
calculation was called the resultant velocity pressure; it is now
more correctly designated the weighted average velocity pressure. It still maintains the symbol of VPr.
Remembering that energy must be conserved at any junction point, the energy entering each of the two airstreams
would be: Q(TP) = Q(SP + VP). The first law of thermodynamics states that the sum the energy in each branch must
equal the energy leaving, or:
Q1(VP1 + SP1) + Q2(VP2 + SP2) = Q3(VP3 + SP3)
+ Losses
[9.19]
In this Manual, branch entry losses are not considered in the
main duct (i.e., F1 = 0) and F2 is provided by Figure 9-f.
Assuming the branch entry in Figure 9-7 is balanced and there
are no dynamic losses due to the fitting (i.e., F2 = 0), there may
still be an additional change in static pressure due to the acceleration or deceleration of the gas stream. The following equation shows this effect:
The static pressure immediately following a junction can be
determined as follows:
SP3 = SP1 – (VP3 – VPr)
[9.22]
With the data shown in Figure 9-7 (IP), determine the static
pressure requirement at point 3.
VPr = [(1,935)(0.79)/(2,275)] + [(340)(0.95)/(2,275)]
= 0.81 "wg
SP3 = SP1 – (VP3 – VPr)
= -2.11 – (1.09 – 0.81)
= -2.11 – 0.28
= -2.39 ''wg
The metric solution using data from Figure 9-7 (SI) is:
SP3 + VP3 = SP1 + (Q1/Q3) VP1
[9.20]
The last two terms on the right of this equation are defined
as the weighted average VP:
VPr = (Q1/Q3) VP1 + (Q2/Q3) VP2
However, if VP3 is greater than VPr, an acceleration has
occurred. When this difference equals or exceeds 0.1 "wg [25
Pa], the difference between the VP downstream of the junction
and the weighted average VP (i.e., VP3 – VPr) is recorded in
Row 37 of the downstream branch where the increase in
velocity is considered. This value represents the necessary loss
in SP required to produce the increase in kinetic energy as air
travels from the branches into the main duct. As such, it must
be accounted for in the static pressure losses of that segment.
EXAMPLE PROBLEM 9-8 (Weighted Average VP)
where the subscripts refer to the ducts shown in Figure 9-7.
+ (Q2/Q3) VP2
When VP3 (i.e., the VP downstream of the junction) is less
than VPr, a deceleration of the airstream has occurred through
the junction resulting in a slight regain in SP. No adjustment is
made in the system design in this case.
[9.18]
Note that the overall losses would be:
Losses = F1Q1VP1 + F2Q2VP2
velocity pressures include the density effects. Also note that, if
the flow rate through one branch was changed to balance at the
branch entry, the corrected velocity pressure (VPcorr) and corrected flow rates (Qcorr) should be used in Equation 9.21. Note
that VPr is not a measurable value in the system.
[9.21]
where:
VPr = weighted average VP of the combined
branches, "wg [Pa]
Q1 = flow rate in branch #1, acfm [am3/s]
Q2 = flow rate in branch #2, acfm [am3/s]
Q3 = combined flow rate leaving the junction,
acfm [am3/s]
Note that the above equation is valid for all conditions,
including merging different density gas streams, as long as the
VPr = [(0.97)(235)/(1.14)] + [(0.17)(271)/(1.14)]
= 240 Pa
SP3 = SP1 – (VP3 – VPr)
= -525 – (325 – 240)
= -525 – (85)
= -610 Pa
Therefore, in this situation, an additional 0.28 "wg [85 Pa]
should be added to the junction SP to account for losses in
pressure due to the acceleration of the airstream.
Local Exhaust Ventilation System Design Calculation Procedures
9-25
removing the effects of the VP at the inlet to the fan (Equation
9.27), yielding the following equation for SSP:
SSP = SPout – SPin – VPin
[9.23]
The values used for calculating SSP are taken from the calc
sheet whereas the values for calculating FSP are based on
manufacturers’ test data. Where these two data points intersect
is the predicted operating point.
FIGURE 9-7 (IP). Branch entry velocity correction
9.10.2 Fan Total Pressure (FTP). Fan total pressure is the
increase in total pressure (TP) through or across the fan and
can be expressed by the equation:
FTP = TPout – TPin
9.10
SYSTEM AND FAN PRESSURE CALCULATIONS
Local exhaust system calculations are based on static pressure; pressures indicate performance of hoods and balancing
or governing pressures at junctions are measures of static pressure. Additionally, the goal of performing system calculations
described in this chapter is to determine the system static pressure (SSP) that can be measured directly in the field as
described in Appendix C.
Most fan rating tables are based on fan static pressure
(FSP). The SSP from the calc sheet is the basis for the determination of the FSP and proper selection of the fan. This section describes the definition of FSP and fan total pressure
(FTP) as provided by the Air Movement and Control
Association (AMCA). Details regarding FSP, FTP and other
terms associated with fan selection are located in Chapter 7.
Both FSP (or FTP) data and SSP (or STP) data to predict system operating points.
9.10.1 System Static Pressure (SSP). System static pres-
sure represents the pressure needed to overcome the losses in
energy as a gas moves through the duct system. It is determined from data on the calculation sheet and is used to specify
the required fan pressure (static or total). To place SSP on the
same graphic representation (fan/system curve; see Section
9.11), the units of measurement must be the same as FSP
(Section 9.10.3). This transposition provides SSP by also
[9.24]
Discussions of TP are provided in Chapter 3, Section 3.6.
Some fan manufacturers base catalog ratings on FTP. To select
a fan on this basis the FTP is calculated noting that TP = SP +
VP:
FTP = (SPout + VPout) – (SPin + VPin)
[9.25]
When VPin = VPout, Equation 9.25 can be simplified to:
FTP = SPout – SPin
[9.26]
9.10.3 Fan Static Pressure (FSP). The AMCA Test Code
defines the FSP as follows:
The static pressure of the fan is the total pressure diminished
by the fan velocity pressure. The fan velocity pressure is
defined as the pressure corresponding to the air velocity at the
fan outlet.(9.1)
Fan static pressure can be expressed by the equation:
FSP = FTP – VPout
[9.27]
FSP = SPout – SPin – VPin
[9.28]
or
Fan static pressure is a term derived from the method of
testing fans and is the value provided by most manufacturers
in their fan selection tables (see Chapter 7). These are not from
the system calculations but empirically derived or computer
generated data for the fan.
NOTE: For the remainder of this chapter, the term fan pressure will apply to both FSP and FTP.
9.10.4 Use of System Static Pressure to Specify a Fan.
The SSP calculation is based on the same formula used to
determine fan static pressure. Therefore, an estimate for the
required FSP can be obtained by determining the SSP and
then:
1) a safety factor,
2) adding any system effect factors (see Chapter 7,
Sections 7.3.2 and 7.4), and
3) addressing provisions for pressure variations (i.e.,
changing ΔP of baghouse during operation).
FIGURE 9-7 (SI). Branch entry velocity correction
In selecting a fan from catalog ratings, the rating tables
should be examined to determine whether they are based on
FSP or FTP. Most centrifugal fans used for LEV systems will
9-26
Industrial Ventilation
be specified using FSP. Fan system effects (see Chapter 7)
should also be considered when selecting a fan. Design appropriate lengths of straight duct entering and leaving centrifugal
fans, as they are especially sensitive to abrupt directional
changes. These fans will require more horsepower and tip
speed if there are elbows or other interferences close to the
fan’s inlet or outlet. The pressure rating can then be calculated
keeping in mind the proper algebraic signs (i.e., VP is always
positive, SPin is usually negative and SPout is usually positive).
A fan curve is developed for a selected fan at a particular
speed. The operating point is an estimated point only because
there could be a change in the SSP as the bag filter pressure or
other values change during operation. Similarly, there may be
multiple fan curves if a variable speed drive and/or fan
dampers are utilized or if fan temperature is changing with the
process. Such conditions could yield varied operating points
and these must be checked to ensure stable operation under all
possible conditions.
The final selection of the fan must also consider the air density. Most fan tables and curves are printed for standard conditions. The final equivalent SSP, calculated and then altered to
meet the above FSP or FTP requirements, must be adjusted for
air density using Equation 7.15 (IP or SI) from Chapter 7:
NOTE: When accounting for system effects (see Chapter 7,
Section 7.4), the fan curve is not altered from the manufacturer’s information. The impact of system effects are considered
in the system calculations as additional system resistance,
which is accounted for in the SSP. This additional resistance
results in a new system curve and intersection point with the
fan curve.
Pe = Pa /df
[9.29]
Note that Equation 9.29 may be used to solve for either the
equivalent FSP or FTP.
The values for df are those shown on the calc sheet in the
segment at the fan inlet. The df at the fan inlet should include
the factor for change in absolute pressure – particularly if the
fan inlet static pressure is below -20 "wg [-5 kPa].
Continuing Sample System Design #1, the SSP and an estimate for FSP can be made from values on the calc sheet. Note
that IP and SI units do not correspond since they have two different inlet conditions. At the outlet of the fan, the SP is +0.3
"wg [+51 Pa]. At the inlet to the fan, the SP is -6.2 "wg [-1470
Pa]. The VP at both locations is 0.51 "wg [116 Pa]. From
Equation 9.23, the system static pressure (SSP) = 0.3 – (-6.2)
– 0.5 = 6.0 "wg [+51 – (-1470) – 116 = 1405 Pa]. This value
is used to specify the required FSP for fan operation. In this
example, the designer would use the SSP as 6.0 "wg [1405 Pa]
and then may choose an FSP for specification of 6.0 "wg, 6.5
"wg or even 7.0 "wg [or similar metric equivalents] – based on
safety factors or other considerations. The specified FSP is the
one selected from the fan tables after Equation 9.28 is completed and the fan is selected. (Assume df = 1.0 in this example.)
9.11
THE SYSTEM AND FAN CURVE RELATIONSHIP
Two curves may be developed that depict the relationship
between actual volumetric flow rate (acfm) [am3/s] and pressure ("wg) [Pa]. The system curve is developed from the SSP
and its relationship with the system flow rate. The fan curve
also shows the relationship between flow rate and pressure,
but is based on data from the fan manufacturer (see Chapter 7,
Figure 7-18). Given that both curves are plotted using variables with the same units of measure, they can both be plotted
on the same graph.
The intersection of the system curve and the manufacturer’s
provided fan curve is the calculated (predicted) operating point
(see Chapter 7). Note that the intersection of fan and system
curves is an approximation.
When fan and system controls are designed for constant airflow operation, some single point of stability may be accomplished. However, most systems are dynamic with changing
flows and pressures as the system’s physical conditions
change. As mentioned above, these conditions may include
changes in: damper settings (manual or automatic), filter differential pressure (ΔP), water flow in a scrubber, or temperature or moisture from a process being ventilated.
Figure 7-21 in Chapter 7 provides an example of two distinct
fan and system curves that may be encountered. In this example, fan curves PQ1 and PQ2 could represent the same fan operating at two different speeds. Similarly, there could be multiple
fan curves indicating different damper settings or temperatures.
System curves A1-A2 and B1-B2 could represent identical systems but with varying pressure conditions within the system.
For example, curve B1-B2 could indicate the operation when
the baghouse filtration media is relatively clean at startup.
Curve A1-A2 could indicate a more restrictive system as the
media becomes laden with dust right before cleaning (resulting
in a higher ΔP). The system curve would then be a family of
curves between the lines indicated by A1-A2 and B1-B2.
If the fan were selected at a constant speed (e.g., PQ2) with
no damper controls, the operating airflow and pressure in the
system could vary between points B2 at start-up and A2 as the
baghouse ΔP increases to a maximum. Please note in Chapter
7 that not all system components operate on the basis of a variance in pressure proportional to square of volume ratio. In particular, the fluctuation in pressure with respect to changes in
bag surface velocity (air/cloth ratio, fpm) [m/s] may be closer
to a linear relationship (Chapter 7, Figure 7-15). Therefore, the
overall system curve may actually have a component that is
not operating as the remainder of the dynamic losses in the
system. If the filter bag losses are a significant contributor to
the total system losses (i.e., more than 50%), the filter manufacturer may need to be consulted to assist in the expected values for changes in ΔP with respect to airflow change in the system.
Local Exhaust Ventilation System Design Calculation Procedures
9.12
SAMPLE SYSTEM DESIGN #2 (MULTI-BRANCH
SYSTEM AT STANDARD AIR CONDITIONS)
Figure 9-8 depicts an LEV system discussing the calculations
for a multi-branch system where the Balance-by-Design Method
is used to equalize static pressures at all junctions. Calculation
sheets shown in Figure 9-12 (IP & SI) illustrate the orderly and
concise arrangement of data and calculations.
The problem considered is a bulk powder handling system.
A minimum transport velocity of 3,500 fpm [18.00 m/s] is
used throughout the problem except after the discharge from
the baghouse where clean air is handled. The system has some
hoods defined in Chapter 13 but Hood 1 requires assumptions
to be made for this special operation. This problem will consider the air at Standard conditions (70 F [21 C], no moisture
and the system at sea level; df = 1). See Section 9.13 for an
example of system design at non-standard air conditions.
The first step to a normal design procedure is either to create
a sketch or single line drawing of the system (Figure 9-9) or
mark up a drawing of the system (Figure 9-10). The sketch
will include the start and end points for each segment
(Segment Identification) eventually to be placed in Row 1 of
the calc sheet.
The operations, hood designations used on the diagram, VSplate references and required flow rates are then presented in
9-27
table format either on a separate sheet or directly on the drawing or sketch. A sample from this problem is shown in Figure
9-11.
With the information from the sketch and Figure 9-11, the
data can be entered to the calc sheet (Figure 9-12). The method
would be to enter information from the top of each column.
Normally, the designer will start with the hood farthest from
the fan and/or with the most junctions between the hood and
the fan. In this case, begin with Hood 1. From the sketch,
Hood 1’s duct combines with Hood 2’s duct at junction “A”.
The segment from Hood 1 would be designated “1-A” for the
start and end point of the segment. This is placed in Cell A/1
of the calc sheet.
The first 14 rows at the top of each column represent the
basic information for a segment and include the flow, duct
size, air conditions (temperature, moisture, etc.) and velocity
pressure. Some of these data are taken from references, such
as the VS plates, while others are calculated.
Beginning in Column A and working down data are input as
follows:
Row 2:
Dry-bulb temperature for standard air is 70 F
[21 C] by definition.
Row 3:
Actual flow rate (acfm) [am3/s] is taken from the
data compiled in Figure 9-11 and includes effects
of density due to elevation, temperature, moisture
and absolute pressure.
Row 4:
Minimum transport velocity is either taken from
Chapter 13 or Table 5-1 in Chapter 5.
Row 5-9: These data are not required because standard air
has no moisture or added heat and df =1.0 (Sample
System Design Problems 3 and 4 will consider
these rows). The row for scfm is used only when
balancing airstreams containing moisture and/or
heat.
FIGURE 9-8. Sample Bulk Powder Handling System –
Sample System Design #2
Row 10:
The target duct area is calculated using the
formula Q = VA and solving for “A”. Flow is
taken from Row 3 and Minimum Transport
Velocity is taken from Row 4.
Row 11:
Selected duct diameter is determined by choosing
the next smaller standard size after calculating
Row 10. For example, in A/10 the calculated duct
size is 0.429 ft2 [0.039 m2]. From Table 9-2 it can
be seen that there is no regular duct size for that
area. The designer would choose either an 8"
[220 mm] diameter duct (area = 0.349 ft2 [0.038
m2]) or a 9" [250 mm] diameter duct (area = 0.442
ft2 [0.049 m2]); because the larger duct will result in
a velocity less than the 3,500 fpm [18 m/s]
specified in Cell A/4, the smaller duct is chosen for
this segment. (Note that IP and SI units will not
match because of duct size selection differences.)
9-28
Industrial Ventilation
FIGURE 9-9. Single line sketch – Sample System Design #2
Row 12:
After the 8" [220 mm] duct is selected its actual
area is inserted from Table 9-2.
Row 13:
The duct velocity is now calculated using the flow
rate in Row 3 and the actual duct area in Row 12.
Row 14:
The duct velocity pressure (VPd) is calculated
from the velocity in Row 12 and Equation 4 from
calc sheet and/or Table 9-4. This VP (1.15 "wg
[204 Pa]) becomes the base that is multiplied by
loss factors for the remainder of the calculations.
FIGURE 9-10. Elevation drawing – Sample System Design #2
of the segment (length of duct, number of
elbows, fitting losses, and any other special
characteristics such as a filter). Note that all of
the information in this section except for the
length of duct and number of elbows are
factors (dimensionless). These values are
totaled in Row 33 and multiplied by the VP in
Row 14. All losses in the segment that are a
function of VP are accumulated and then
multiplied by the VP in that segment to get the
losses in "wg [Pa].
The remainder of the column is then calculated based on the
physical conditions of the system. The important data requested and input into the sheet include:
Row 34:
These are the duct component losses for the
column and are stated in "wg [Pa].
Rows 15-19:
Data required for a slotted hood (see Chapter
6); this may include slot area, slot velocity
and slot loss factor. (Not required for
the segment in Column A.) This information
may also be used to calculate the static
pressure loss for an orifice.
Row 35:
A cell where additional losses in "wg [Pa] can be
placed (e.g., the DP across a filter, spark arrester,
etc.).
Row 36:
The cell where the Weighted Average Velocity
Pressure (VPr) is calculated at a junction (see
Section 9.9).
Rows 20-24:
Required for all hoods (except for orifices)
with or without slots and includes the physical
characteristics and shape factors for the hood.
This information comes either from Chapter 6
or the VS plates located in Chapter 13.
Row 37:
Rows 25-33:
Data considering all of the physical aspects
If the velocity increases in a junction so that
the downstream VP is higher than the value
in Row 36, then the difference must be
added in this cell; in effect the VPr must be
less than the VP of the upstream junction. It
is good practice to calculate VPr at every
Local Exhaust Ventilation System Design Calculation Procedures
9-29
FIGURE 9-11. Basic system information – Sample System Design #2
Row 38:
junction (see Section 9.9). Note that values
for VP are always positive.
is selected there will be more air than desired pulled through
Hood 1.
This row represents the summation of all
losses in this segment. In the case of
Column A in this example, it states that if
-1.8 "wg [-318 Pa] of pressure is applied at
junction A, then the duct and hood system
from Hood 1 will exhaust 1,500 acfm [0.70
am3/s]. If more negative pressure is applied,
then more air will flow, etc. The key to a good
design is to get the proper pressure at that
junction.
Per Section 9.8.1, the ratio of the SP values for branches 1-A
and 2-A are calculated:
Additional notes:
Cells A/3 and A/8: Since the density factor is 1.0 (standard
air) in this example, the acfm = scfm. In this case, the values
in Rows 3 and 8 for all branches are equal. When designing
systems with standard air only, the values in Row 8 can be left
out of the calc sheet and all calculations are done with acfm.
Cells A/38 and B/8: This is the classic example of a system
balance issue (see Section 9.8). The calc sheet states that -1.8
"wg [318 Pa] of pressure will deliver 1,500 acfm [0.70 am3/s]
from Hood 1, but -3.1 "wg [683 Pa] of pressure is needed at
the same junction to pull the 200 acfm [0.10 am3/s] from Hood
2. There can only be one value for SP at the junction. If the
branch whose SP is lowest in magnitude (-1.8 "wg [318 Pa]) is
selected then there will not be enough energy to pull the 200
acfm [0.10 am3/s] from Hood 2. At the same time, if the branch
whose SP is greatest in magnitude (i.e., the governing branch)
The branch with the lower resistance should be redesigned
since the SP ratio is higher than 1.2. This is accomplished in
the third column of the calc sheet and its segment designated
1'-A. In this case, a smaller duct (i.e., decreased size from 8"
to 7" diameter) [In SI units the size is reduced from 220 mm
to 190 mm] is selected, which increases the velocity in the duct
segment from Hood 1. This increase in velocity increases the
friction and, therefore, the resistance in the segment. When the
new column is completed the required SP is now -3.2 "wg in
the IP solution [-606 Pa in the SI solution]. Now segment 1'-A
is the governing branch because its SP is greater in magnitude
than the SP in segment 2-A. The SP ratio is again tested:
FIGURE 9-12 (IP). Velocity Pressure Method Calculation Sheet – Sample System Design #2
9-30
Industrial Ventilation
FIGURE 9-12 (SI). Velocity Pressure Method Calculation Sheet – Sample System Design #2
Local Exhaust Ventilation System Design Calculation Procedures
9-31
9-32
Industrial Ventilation
The SP ratio now falls below the 1.2 value required for segment redesign. However, the SP in Hood 2 [Hood 1] still does
not equal that of Hood 1 [Hood 2].
[Qcorr = 0.70(-683/-606)0.5]
This value is entered in Row 42. Next, the velocity in the
duct and VP are corrected and entered into Rows 42 and 43,
respectively.
Cell D/3: The new airflow of 1,707 acfm [0.84 am3/s]
required in this segment is the sum of the flow from 2-A (207
acfm) [0.10 am3/s] plus the value in 1'-A (1,500 acfm) [0.74
am3/s].
NOTE: This is one of the potential disadvantages listed in
Chapter 4, Table 4-1 for the Balance-by-Design Method. In
place of the design 1,700 acfm [0.80 am3/s] of (1,500 + 200)
[0.70 + 0.1] originally intended, this method now results in a
recalculated design flow of 1,707 acfm [0.84 am3/s]. While the
increase in airflow is small, it will result in an increase in
power consumption at the fan. The use of blast gates here
would have resulted in a savings of airflow, pressure and
horsepower.
Cells D/36 and F/36 (IP only): These show two possible situations when considering weighted average velocity pressure.
The value for the weighted average velocity pressure of the
branches entering Junction B (see Section 9.9) is computed
using the values in D/3 (Q1), E/41 (Q2), F/3 (Q3), D/14 (VP1),
and E/43 (VP2). From Equation 10 on the calc sheet:
This value is inserted in F/36 and compared to the VP in the
next duct segment (B-C). Since VPr is less than the VP calculated in the B-C segment, the difference between the two values (1.10 – 0.99 = 0.11 "wg) is added to the losses for the B-C
segment and shown separately in F/37.
However, the weighted average velocity pressure at
Junction A shows when the effects can be ignored. In that
case, the VPr for the branches 2-A and 1'-A is calculated with
the same equation to be 1.87 "wg. The VP in the next segment
after the combination (A-B) is 0.93 "wg (shown in D/14).
Since this value is lower than the VPr, air has basically slowed
as it goes through the fitting so there is no added
resistance/loss.
Cells E/38 and E/39 (IP only): Note that the governing static
pressure at Junction “B” is -3.4 "wg. However, the SP requirement for 3-B is only -2.5 "wg. If a test is performed at that
junction the ratio would be:
This would normally require a duct size change in branch
3-B, perhaps a reduction to 4.5". However, the designer ran
the calculation and determined that this new pressure would
now be governing and force even more volume from branch
A-B. These types of decisions can be encountered during the
system design. Rather than redesign the smaller duct, the 1.4
ratio was applied to Equation 9.17 and a new volume (593
acfm) was calculated for branch 3-B.
Cell J/20: The baghouse in this case was specified with a
maximum pressure drop across the filter media of 6.0 "wg
[1500 Pa]. This is recorded as Other Losses in Cell J/35.
Since the baghouse loss is not ‘flange-to-flange’ there are other
losses as the air is turbulent through the baghouse and adds
resistance to the system. If the manufacturer does not provide
the information for these losses, then a normal assumption is
to use the losses for a flanged duct end (Figure 9-a). That
would allow for 0.60 VP of loss (added in Cell J/20) and
another increase in velocity from the extremely low speed
(usually 3.0 to 10.0 fpm [0.02 to 0.05 m/s]) through the filter
media and reaccelerated to 3,165 fpm [16.04 m/s] in segment
C-E. This increase in velocity requires the same consideration
as the energy exchange in a hood, i.e., 1.0 VP. (See discussion
of this coefficient in Chapters 3 and 6.) This is added in Cell
J/22 and the baghouse losses are treated similarly to the total
losses in a hood.
Cell K/30: Note that there is no added loss for the ‘no-loss’
stack as shown in the sketch (Figure 9-9). If a rain cap had
been used then more resistance would have been added in this
section and more horsepower would be required.
Cells J/14, J/40, and K/40: These are the values used to
determine the SSP (Equation 11 on the calc sheet; or Equation
9.23 in Section 9.10.1). When the value for SSP is determined,
the designer can then select an FSP for specification of the fan.
In this case, the value for SSP is (+0.1" – (-11.5") – 0.6") or
11.0 "wg [2868 Pa]. This could be rounded up to 11.0 "wg
[2900 Pa] or even higher. The selected FSP may include some
factor of safety and rounding of values. In this case, it could be
selected at 12.0" [3000 Pa] or some other value. After selection
and review of the fan and system curves, the fan selection may
be changed if it does not appear to be selected for a stable operating condition.
Cell J/3: This is the value for airflow used to select the fan.
Since the air is at standard conditions, the fan would be specified:
2,918 acfm @ 12" FSP @ standard conditions.
Note in the IP example that the FSP includes the small fac-
Local Exhaust Ventilation System Design Calculation Procedures
tor of safety increase from the SSP calculated at 11.0 "wg. [In
the SI example, the fan would be specified at 1.46 am3/s @
3000 Pa @ STP]. The selection of 3000 Pa for the fan in the
SI example includes a factor of safety above the 2868 Pa calculated. This value for FSP must reflect the maximum pressure
drop to be encountered by the filter bags. When the system is
first started there may not be 6 "wg [1500 Pa] of resistance
(DP) across the bags. In that case, the fan will operate at higher
airflow than the design and can cause premature plugging of
the filter media. A volume control damper or variable frequency drive should be considered to keep the system from operating at a volume and power in excess of design.
9.13
CALCULATION METHODS AND NON-STANDARD
AIR DENSITY
The example shown in Sample System Design # 2 (Section
9.12) considers standard air density — something that rarely
occurs in real system design. It simplifies the calculations by
assuming that air is constantly at standard conditions (0.075
lbm da/ft3 and no moisture) [1.204 kg da/m3]. Even though the
effects of moisture, elevation, pressure and temperature can be
small when considered independently, they can have significant, additive effects when considered together.
Fan tables assume standard air density that corresponds to
sea level elevation, no moisture, and 70 F [21 C]. Changes in
air density can come from several factors, including elevation,
temperature, internal duct pressure, changes in apparent
molecular weight (moisture content, gas stream constituents,
etc.), and amount of suspended particulate. The change in air
density must be considered when calculating flow and pressure requirements.
Density factors for different temperatures and elevations are
listed in Table 9-7 (IP and SI). Internal duct pressures will also
change air density and can have a significant effect, especially
at the fan inlet. If there is excessive moisture in the airstream,
the density will decrease.
Suspended particulate is assumed to be only a trace impurity in industrial exhaust systems. If there are significant quantities of particulate in the duct system (> 20 grains/dscf [0.05
kg/m3]), the addition to the airstream density should be
addressed. This field is called material conveying and is
beyond the scope of this Manual. Note that at 20 grains/dscf
[0.05 kg/m3] the particulate represents less than 0.4% of the air
mass rate – significant amounts of air to move a small amount
of particulate. In cases where there is more than 20 grains/dscf
[0.05 kg/m3] of material in the airstream, a factor can be
applied to the losses in this Manual for ‘clean air’. This factor
calculates as:
[9.30]
9-33
The density variation equations of Chapter 3, Section 3.5
demonstrate that, for a constant mass flow rate, an increase in
temperature or a reduction in absolute pressure will increase
the actual flow. It is helpful to remember that a fan connected
to a given system will exhaust the same volumetric flow rate
regardless of air density. The mass of air moved, however, will
be a function of its density.
Additionally, knowing dry-bulb temperature (T) and moisture content (ω), the value of enthalpy (h) can be calculated
approximately from the following equation:
h = 0.24T + ω(1,061 + 0.444T)
[9.31] IP
[h = 1.006T + ω(2501 + 1.86T)]
[9.31] SI
Use of this equation can eliminate errors sometimes occurring from difficulty in accurately reading a Psychrometric
chart.
In cases where two airstreams combine there can also be
cases where moisture is added to an airstream. Section 9.4.3
considers the combining of airstreams where little moisture is
present. Industrial ventilation systems often combine a hot
moist stream with a cooler dry mass. In some cases, the combination can encourage condensation of the moisture from the
hot stream and can be a problem for the design (i.e., condensed
moisture combining with dry dust can plug filters and coat the
duct components). It is important to be able to predict moisture
and heat conditions for these types of combinations.
9.14
SAMPLE SYSTEM DESIGN #3 (SINGLE-BRANCH
SYSTEM AT NON-STANDARD AIR CONDITIONS)
(IP Units Only)
The example shown in Figure 9-13 illustrates the effects of
elevation, moisture and temperature on LEV system design for
these systems. A calc sheet showing the system calculations is
provided in Figure 9-14.
Given: The exit flow rate from a 60" H 24" dryer is 16,000
scfm plus removed moisture. The plant is located at 575 feet
ASL and the dryer exhaust air temperature is 500 F. The dryer
delivers 60 tons/hr of dried material with capacity to remove
5% moisture. Required suction at the dryer hood is -2.0 "wg;
minimum transport velocity is 4,000 fpm.
It has been determined that the air pollution control system
should include a cyclone for dry product recovery and a highenergy wet collector. These devices have the following operating characteristics:
•
Cyclone: Pressure loss is 4.5 "wg at a design flow rate
of 35,000 scfm.
•
High-Energy Wet Scrubber: The manufacturer has
determined that a pressure loss of 20 "wg is required in
order to meet existing air pollution regulations and has
sized the collector accordingly. The humidifying efficiency of the wet collector is 90%.
NOTE: As a practical matter, a high energy scrubber as
9-34
Industrial Ventilation
FIGURE 9-13. System layout – Sample System Design #3
described in this example could have essentially 100% humidifying efficiency. The assumption of 90% humidifying efficiency along with a high pressure drop allows discussion of multiple design considerations in one example and, therefore, has
been selected for instructional purposes.
•
Fan: A size #34 “XYZ” fan with the performance
shown in Figure 9-15 has been recommended.
= 1,200 lbm da/min
Step 1C: Knowing the water-to-dry air ratio and the temperature of the combination, it is possible to determine other qualities of the air-to-water combination. This can be accomplished by the use of the Psychrometric charts (Figures 9-h
through 9-n) that are useful tools when working with humid
air.
REQUIRED:
 = 106.7/1,200 = 0.089 lbm-H2O/lbm da
Size the duct and select fan RPM and motor size.
Dry-Bulb temperature = 500 F (given)
SOLUTION:
Step 1: Find the actual gas flow rate that must be exhausted
from the dryer. This flow rate must include both the air used
for drying and the water removed from the product. Therefore,
the flow rate must be corrected from standard air conditions to
reflect the actual elevation, moisture, temperature and pressures that exist in the duct.
The intersection of the 500 F Dry-Bulb temperature line and
the 0.089 lbm H2O/lbm da line can be located on the
Psychrometric chart (see Point #1 on Figure 9-16). This point
defines the quality of the air and water combination. Other
data relative to this specific mixture can be read as follows:
Dew Point Temperature: 122 F
Step 1A: Find the amount (weight) of water vapor exhausted.
Wet-Bulb Temperature: 144 F
Dryer Discharge = 60 tons/hr of dried material (given)
Humid Volume, ft3 comb/lbm da: 27.5 ft3/lbm da
Since the dryer has the capacity to remove 5% moisture, the
dryer discharge is 95% H dryer feed rate.
60 tons/hr dried material = (0.95) H (dryer feed)
dryer feed =
= 63.2 tons/hr
Moisture removed = (feed rate) – (discharge rate)
= 63.2 tons/hr – 60 tons/hr
= 6,400 lbm/hr or 106.7 lbm/min
Step 1B: Find the amount (weight) of dry air exhausted.
Dry air exhausted = 16,000 scfm at 70 F and
29.92 "Hg (0.075 lbm/ft3 density)
Exhaust rate = (16,000 scfm)(0.075 lbm/ft3)
Enthalpy, BTU/lbm da: 234 BTU/lbm da
Density Factor, df: 0.53
The system is designed at an elevation of 575 feet ASL; this
alters the df further to a value of 0.52. The density factor, drybulb temperature, mass of air and water, scfm and enthalpy are
entered in the appropriate lines on the calc sheet.
Step 2: Proceed with the system design using the calculation
methods from previous Sample System Design #1 and #2.
FIGURE 9-14 (IP). Velocity Pressure Method Calculation Sheet – Sample System Design #3
Local Exhaust Ventilation System Design Calculation Procedures
9-35
9-36
Industrial Ventilation
FIGURE 9-15. Fan rating table
When considering the loss through the cyclone (A-B), the
value is inserted in Row 35. The manufacturer provides the
pressure loss (DP) of the cyclone. In this example, the cyclone
pressure loss is 4.5 "wg at a rated flow of 35,000 scfm. The
pressure loss through a cyclone usually varies as the square of
the change in flow rate and directly with the change in density.
Therefore, the actual loss through the cyclone would be:
and the static pressure at the cyclone outlet would be -4.3 "wg.
There are no reacceleration losses.
The scrubber equipment manufacturer should provide the
information for calculation of changes in flow rate and pressure drop across the wet collector, etc. An important characteristic of wet collectors is their ability to humidify a gas stream.
The humidification process is generally assumed to be adiabatic (i.e., without gain or loss of heat to the surroundings). Water
vapor is added to the combination, but the enthalpy, expressed
in BTU/lbm da, remains unchanged. During the process of
humidification, the point on the Psychrometric chart that
defines the quality of the combination moves to the left, along
a line of constant enthalpy, toward saturation.
All wet collectors do not have the same ability to humidify.
If a collector is capable of taking an airstream to complete adiabatic saturation, it is said to have a Humidifying Efficiency of
100%. The humidifying efficiency of a given device may be
expressed by either of the following equations:
[9.32]
where: hn = humidifying efficiency, %
Ti = Dry-Bulb temperature at collector inlet, F
To = Dry-Bulb temperature at collector outlet, F
Ts = adiabatic saturation temperature, F
or
where: i = moisture content in lbm H2O/lbm da at inlet
o = moisture content in lbm H2O/lbm da at
outlet
s = moisture content in lbm H2O/lbm da at
adiabatic saturation conditions
Note: These formulae also apply in the SI system using temperature in C and ω in kg H2O/kg da.
The designer must find the quality of the air-to-water combination at Point 2, the collector outlet. Humidifying
Efficiency = 90% (given). Dry-Bulb Temperature at Collector
Local Exhaust Ventilation System Design Calculation Procedures
9-37
FIGURE 9-16. Psychrometric chart for humid air – excerpted from Figure 9-j (IP)
Inlet = 500 F (given). Adiabatic saturation temperature (i.e.,
Wet Bulb) = 144 F from inspection of Psychrometric chart.
calculated.
Note: Water content in air is now (1,200 lbm da/min)(0.16
lbm H2O/lbm da)
= 192 lbm H2O/min
thus:
To = 180 F
Q = 20.5 ft3/lbm da × 1,200 lbm da/min
Therefore, the air leaving the collector will have a dry-bulb
temperature of 180 F and an enthalpy of 234 BTU/lbm da as
the humidifying process does not change the total heat or
enthalpy.
The point of intersection of 180 F dry-bulb and 234
BTU/lbm da on the Psychrometric chart (see Point #2 on
Figure 9-16) defines the quality of the air leaving the collector
and allows other data to be read from the chart as follows:
Dew Point Temperature
143 F
Wet-Bulb Temperature
144 F
3
Humid Volume, ft /lbm da
20.5 ft3/lbm da
Enthalpy, BTU/lbm da
234 BTU/lbm da
Density factor, df
0.76

0.16
The density factor is recalculated at 0.74 to consider elevation. Required information is placed in the calc sheet and,
using the humid volume, the acfm going into the scrubber is
= 24,600 acfm
The scrubber loss was stated to be 20 "wg, so the static pressure at the wet collector outlet would be -24.3 "wg.
Step 3: Previously, in low-pressure LEV systems, (where
the negative pressure at the fan inlet was less than -20 "wg),
the effect of the negative pressure on airstream density was
usually ignored (i.e., its effect was less than 5%). In practical
system design, the other factors that affect density (temperature, moisture, elevation) can be additive so that the inlet pressure can be significant when specifying the fan. Systems
designed at air temperatures less than 100 F and near sea level
(df = 1) can still ignore fan inlet pressure if the values are
between +20 and -20 "wg. However, as the pressures decrease
(magnitude of negative pressures increase), it is understood
that gases expand to occupy a larger volume. Unless this larger
volume is anticipated and the fan is sized to handle the larger
flow rate, it will have the effect of reducing the amount of air
that is pulled into the hood at the front end of the system.
From the energy equation for flow in a duct without heat
9-38
Industrial Ventilation
transfer (see Chapter 3 and Section 9.5.3):
acfm. Since the fan selected has a 34-inch diameter inlet (area
= 6.307 ft2), it is convenient to make the duct from the wet collector to the fan a 34-inch diameter.
After the system calculation has been completed, the actual
FSP can be calculated.
FSPa = SPout – SPin – VPin
= +0.1 – (-24.3) – 0.8
Considering the Ideal Gas Equation, this would yield:
= 23.6 "wg
Step 5: The equivalent fan static pressure (i.e., the pressure
used to select the fan) is determined by dividing the actual fan
static pressure by the density factor at the fan inlet (see
Equation 9.29). This is necessary since all fan rating tables are
based on standard air.
Up to this point, the air has been considered to be at standard atmospheric pressure, which is 407 "wg. The pressure
within the duct at Point F is -24.3 "wg (i.e., negative only in
relation to the pressure outside the duct which is 407 "wg).
Therefore, the absolute pressure within the duct is 407 "wg
-24.3 "wg = 382.7 "wg.
Q2, the value at the fan inlet = 26,162 acfm
Note: If using Equations 9.11 and 9.14, values may vary
slightly from psychrometric chart.
In this case, the design considers no safety factor and the
actual value of 33.7 is used to select the fan.
Step 6: Interpolating the fan rating table (Figure 9-15) for
26,162 acfm at 33.7 "wg yields a fan speed of 1,562 RPM at
217 BHP.
Since actual density is less than standard air density (and
conveying air with less mass will require less work/energy),
the actual power requirement (PWRa) is determined by multiplying by the density factor, or (217 BHP)(0.70) = 152 BHP.
If a damper is installed in the duct to prevent overloading of
the motor, a 200 HP motor should be selected to accommodate
cold starts (see Chapter 7, Section 7.3.8).
Additional Notes for Sample System Design #3:
Step 4: Absolute pressure also affects the density of the air.
From PQ = ṁRT, the relationship
can be derived. Assuming no heat transfer or change in temperature, the density factor is directly proportional to the density and the equation can be rewritten
If the pressure in the duct is compared to the absolute pressure at standard conditions (407 "wg), this can be calculated:
Cell A/14: Note that the value for VP has already been corrected for density using Equation 4 on the calc sheet. It will
represent a value close to the real check number when performing a balance. The calc sheet is not the most accurate template for predicting actual field conditions. For one thing, the
density would have to be exactly as calculated. If moisture levels or temperatures are different than calculated, this will affect
these values. However, the 0.59 "wg is a good starting point to
check airflows and conditions when commissioning the system and attempting to meet flow requirements.
Cell B/35: Care must always be taken when entering special
losses and/or loss factors. In this case, the loss through the
cyclone was calculated in "wg (see Step 2, Sample System
Design #3). Some equipment may be rated with losses in value
of velocity pressures (i.e., 2.0 VP). In those cases, the loss factor
would be added in Row 30 (Special Fitting Loss Factor) instead
of Row 35.
df2 = 0.70.
This is now the ‘real’ value for density factor used for the
fan specification. It considers temperature, moisture, elevation
and now absolute pressure in the duct.
9.15
SAMPLE SYSTEM DESIGN #4 (ADDING A
BRANCH TO AN EXISTING SYSTEM AT NONSTANDARD AIR CONDITIONS) (IP Units Only)
The duct from the wet collector to the fan can now be sized.
The actual flow rate leaving the wet collector was 26,162
A second example is included where a new hood connection
is added to the original duct system as an afterthought (Figure
Local Exhaust Ventilation System Design Calculation Procedures
9-17). This is not good practice under almost any circumstances. The original design is always compromised and there
can be cases where material will settle in the duct, airflow will
be reduced to other connections, and/or system changes in
flow or pressure will cause the fan to operate in an unstable
manner. If the addition of one or more ducts is made, the system calculation principles still apply. Being mindful of maintaining minimum transport velocities, losses can be calculated
for the added flow at the fan. The following example should
not in any way be considered an endorsement of this practice.
It is included only to show that calculations and system adjustments can be made to get the system into balance (if suitable
resources are available in the duct, fan, motor, collection
device and electric power source).
In this case, a hood similar to the bagging hood shown in
VS-15-02 is connected through a properly sized branch and
tapped into the 38" diameter duct coming from the dryer.
When the decision is made to proceed on this basis, many factors must be considered:
1) Combining hot and moist airstreams with cold air can
cause condensation in duct or collectors. Under normal
conditions, the dry-bulb (DB) temperature should be at
least 35 F [20 C] above the dew point and preferably 50
F [28 C]. The system must also consider start-up and
shut down when the system is especially susceptible to
condensation or the formation of other chemicals such
as acids or alkalies.
2) Downstream velocities may be high enough to cause
premature wear of duct and other parts.
3) Sufficient airflow and transport velocities must be
maintained through all duct system components and at
all hoods.
A new calc sheet (Figure 9-18) shows the alterations that
must be made. A new sketch inserting a new branch duct with a
FIGURE 9-17. System layout – Sample System Design #4
9-39
new segment identification is made. The bagging station is 60'
away from the 38" main duct with (1) 90°, 4-piece, 1.5 r/d
elbow and requires 1,500 acfm. The designer in this case has
chosen to keep all duct the same size (i.e., no size increase in the
main duct between the dryer and the cyclone). All calculations
are done in the same manner as previous examples except a calculation must also be done to allow for the combining of the
ambient air from the bagging station with the hot moist air from
the dryer.
Knowing that mass and energy must be conserved, the conditions downstream of the fitting can be calculated using
Equation 9.14. The mass of dry air (Row 6) and enthalpy (Row
9) are known for the two branches and the mass is known for
the downstream duct since it is simply the sum of the two
branches.
(1200 H 234)1 + (110.3 H 17)2 = (1310.3 H h)3
h3 = 216 BTU/lbm da
The mass of air and water downstream will be the summation of the two values from the new hood and the dryer (Rows
5 and 6 on the calc sheet). Using this information, the conditions in duct 2-A can be determined from the Psychrometric
chart as:
Wet Bulb Temperature
142 F
Dry-Bulb Temperature
471 F*
Density factor, df
0.55

0.081
*Calculated with equation 9.27 [IP]
FIGURE 9-18 (IP). Velocity Pressure Method Calculation Sheet – Sample System Design #4
9-40
Industrial Ventilation
Local Exhaust Ventilation System Design Calculation Procedures
9-41
The density factor is again recalculated at 0.54 to consider
elevation, resulting in a flow rate of 34,972 acfm, the required
information is placed in the calc sheet.
An air bleed (Figure 9-19). From Figure 9-a, he = 1.78 VP.
Use the following steps to design an air bleed/orifice plate.
Note that the static pressure requirement for the new branch
is 4.0 "wg at junction A and the requirement at the same junction for the dryer is only -2.2 "wg. Normally there would be a
change in duct design or selection of new airflows to increase
the flow to the dryer and balance the pressure from each branch.
In this case, the dryer is sensitive to the static pressure from the
duct and cannot be altered. This is a good example of where
dampers or orifice plates can be used to balance the system.
2) Determine airflow rate in main duct according to
design or future capacity, or directly from temperature
or moisture considerations.
The remainder of the calculation process is identical to the
first example and airflow at the fan inlet is now required to be
28,950 acfm. After the system calculation has been completed,
the new system conditions can be determined:
SSP = SPout – SPin – VPin
= +0.1 – (-26.5) – 0.9"
= 25.7 "wg
The fan will now be required to operate at increased airflow
and pressure to meet the design requirements but with a significant increase in horsepower. The fan speed is recalculated at
1,602 RPM and the horsepower required under cold conditions is now 230.
The fan will now need to operate at increased airflow and
pressure, but if a 200 HP motor was originally selected, it will
not be large enough for a cold start-up of the modified system
(see Chapter 7, Section 7.3.8). In that case, design and/or hardware changes will have to be made to damper the fan at startup until sufficient heat is in the system to reduce the power
requirements.
NOTE: For Cell A/40: The static pressure required to deliver
33,508 acfm from segment 1-A is -2.2 "wg. Since the system is
not being balanced by design, the determining value at Junction
A (-4.0 "wg from segment 2-A) is not used to recalculate the
conditions in 1-A. Instead, a blast gate, orifice, or other damper
will be used to balance the system. The fan must be able to
deliver 4.0" at this junction to pull all of the air designed for the
bagging station. The difference between the junction determining pressure (4.0 "wg) and the losses in the duct segment from
1-A (-2.2 "wg) is the added amount of loss that will be required
from the blast gate or orifice plate (1.8 "wg).
1) Calculate SP for branch duct to junction (A).
3) Qair bleed = (Q1-A) – (Q2-A)
4) SPair bleed = │SP1-A│
= calculated SP of air bleed branch
= │SPh│+ Duct Component Losses
5) Duct Component Losses = [(Fd) + (# 90° elbows)
(Fel) + (Fen)] × VPd
6) SPh = │SP1-A│– Duct Component Losses
7) VPair bleed = │SPh│/(1.78 + 1)
8) Vair bleed: from VPair bleed and Table 9-2 or 9-3
9) Aair bleed = (Qair bleed/(Vair bleed)
EXAMPLE PROBLEM 9-9 (Air Bleed to Reduce Duct
Temperature)
The melting furnace ventilated in Example Problems 9-2–9-5
has a temperature of 196 F [91 C]. An air bleed must be added
to reduce the temperature to 125 F [52 C] for entry into the baghouse. The outside air temperature will be 70 F [21 C]. The air
bleed will be placed into the duct system where accumulated
losses from the furnace hood equal 3.2 "wg [800 Pa]. The total
duct component losses in the air bleed duct segment equal 1
"wg [250 Pa]. Calculate the size of the air bleed orifice.
From Section 3.9 in Chapter 3:
(ṁa)(Ta) + (ṁf)(Tf) = (ṁcomb)(Tcomb)
(ṁa)(460 + 70) + (720.3)(460 + 196) = (720.3 + ṁa)
9.16 AIR BLEED DESIGN
Air bleeds are used at the ends of branch ducts to provide
additional airflow to transport heavy material loads or at the
ends of a main duct to maintain minimum transport velocity.
Other designs use air bleeds to introduce additional air to
reduce air temperature and/or to assist in balancing the system
(e.g., where a branch has been removed).
FIGURE 9-19. Air bleed opening
9-42
Industrial Ventilation
(460 + 125)
REFERENCES
[(ṁa)(273 + 21) + (5.37)(273 + 91) = (5.37 + ṁa)
(273 + 52)]
9.1
Air Movement and Control Association, Inc.: AMCA
Standard 210-74. Arlington Heights, IL.
9.2
E. Ravert: Private Communication to G. Lanham
(November, 2012).
Solving for the mass flow rate of the air bleed:
ṁa = 930 lbm/min [6.76 kg/s]
Solving for Qstd:
Qstd = (ṁa)/(rstd) = (930 lbm/min)/(0.075 lbm/ft3)
= 12,400 scfm @ 70 F
APPENDIX A9 PRESSURE MEASUREMENT IN THE SI
SYSTEM
[Qstd = (ṁa)/(rstd) = (6.76 kg/s)/(1.204 kg/m3)
= 5.61 nm3/s @ 21 C]
In the research for this Manual, it was noted that within the
SI system there are different units used worldwide for pressure. The classic measurement for pressure in the SI system (in
ranges for ventilation systems) is the Pascal (Pa). A soft conversion (see Foreword and Definitions) from inches of water
("wg) to the Pascals is 249.1 Pa per "wg. In this Manual, the
hard conversion of 250 Pa per "wg was used, a value less than
0.0025 "wg from the soft (249.1) conversion. This can also be
stated as 0.25 kilo-Pascals (kPa).
SPair bleed = 3.2 "wg
= │SPh│ + Duct Component Losses
= │SPh│ + 1 "wg
[SPair bleed = 800 Pa = │SPh│ + 250 Pa]
│SPh│ = 2.2 "wg [550 Pa] = (1.78 + 1)VPair bleed
VPair bleed = (2.2 "wg)/2.78 = 0.79 "wg
[VPair bleed = (550 Pa)/2.78 = 198 Pa]
Vair bleed = 3,560 fpm; from Table 9.2 (IP)
[Vair bleed = 18.15 m/s; from Table 9.2 (SI)]
Aair bleed = (Qair bleed/(Vair bleed)
= (12,400 scfm)/(3,560 fpm) = 3.48 ft2
[Aair bleed = (5.61 nm3/s)/(18.15 m/s) = 0.31 m2]
Unlike a duct size, this would be the actual size of the circular orifice opening in the air bleed (i.e., 25.3" diameter).
Note that the new airflow required for specification of the
baghouse and fan at 4,300′ [1300 m] ASL and 125 F [52 C] is
28,117 acfm [12.87 am3/s]. This example is for “dry” air only. If
there is moisture present, then enthalpy will need to be used as
per Sample System Design #3.
Please note that if calculations are done using a computer
(spreadsheet program, etc.), do not round numbers until the
final computation at the fan and discharge points.
An alternative measurement method is millimeters of water
(mmwg). This can also be referred to as millimeters water
gauge. Using a U-Tube manometer yields a direct measurement
of "wg. If a similar measurement is taken in the SI system, the
values would be measured directly in mmwg. This becomes a
conversion where (soft conversion) one inch = 25.4 mm.
Comparing the hard conversion of the two methods for 1
"wg (250 Pa and 25 mmwg) there is less than 1.5% from the
soft conversion values. In this case 10 Pa is considered equal
to 1 mmwg.
This Manual utilizes Pa for pressure measurement in the SI
system. Those using mmwg for pressure measurement need
only to divide pressure in Pa by 10 to achieve a similar value
in mmwg and calculate pressure density factor (dfp) from:
dfp = (10,338 + SP) / (10,338)
Those needing or desiring to do so can calculate velocity or
velocity pressure with VP in units of mmwg from:
V = 4.043(VP/df)0.5; VP = df(V/4.043)2
The remaining SI values, airflow, velocity, dimensions, etc.
are valid for both the Pa and mmwg methods.
Local Exhaust Ventilation System Design Calculation Procedures
TABLE 9-2. Area and Circumference of Circles
9-43
9-44
Industrial Ventilation
TABLE 9-3 (IP). Velocity Pressure to Velocity Conversion — Standard Air
Local Exhaust Ventilation System Design Calculation Procedures
TABLE 9-3 (SI). Velocity Pressure to Velocity Conversion — Standard Air
9-45
9-46
Industrial Ventilation
TABLE 9-4 (IP). Velocity to Velocity Pressure Conversion — Standard Air
Local Exhaust Ventilation System Design Calculation Procedures
TABLE 9-4 (SI). Velocity to Velocity Pressure Conversion — Standard Air
9-47
9-48
Industrial Ventilation
TABLE 9-5 (IP). Duct Friction Loss Factors per Foot of Duct Length, F'd
Local Exhaust Ventilation System Design Calculation Procedures
TABLE 9-5 (IP) (Cont.). Duct Friction Loss Factors per Foot of Duct Length, F'd
9-49
9-50
Industrial Ventilation
TABLE 9-5 (SI). Duct Friction Loss Factors per Meter of Duct Length, F'd
Local Exhaust Ventilation System Design Calculation Procedures
TABLE 9-5 (SI) (Cont.). Duct Friction Loss Factors per Meter of Duct Length, F'd
9-51
TABLE 9-6 (IP). Circular Equivalents of Rectangular Ducts (in)
9-52
Industrial Ventilation
TABLE 9-6 (SI). Circular Equivalents of Rectangular Ducts (mm)
Local Exhaust Ventilation System Design Calculation Procedures
9-53
9-54
Industrial Ventilation
TABLE 9-7 (IP). Air Density Correction Factor (Temperature and Elevation Only), dfT H dfe
TABLE 9-7 (SI). Air Density Correction Factor (Temperature and Elevation Only), dfT H dfe
Local Exhaust Ventilation System Design Calculation Procedures
9-55
9-56
Industrial Ventilation
FIGURE 9-b (IP). Friction chart for sheet metal & plastic ducts (equivalent sand grain roughness height = 0.00015 feet)
Local Exhaust Ventilation System Design Calculation Procedures
9-57
FIGURE 9-c (IP). Friction chart for sheet metal & plastic ducts (equivalent sand grain roughness height = 0.00015 feet)
9-58
Industrial Ventilation
Local Exhaust Ventilation System Design Calculation Procedures
9-59
9-60
Industrial Ventilation
Local Exhaust Ventilation System Design Calculation Procedures
9-61
9-62
Industrial Ventilation
FIGURE 9-h (IP). Psychrometric chart — 30 F to 115 F Dry Bulb Temperature
Local Exhaust Ventilation System Design Calculation Procedures
FIGURE 9-i (IP). Psychrometric chart — 60 F to 250 F Dry Bulb Temperature
9-63
9-64
Industrial Ventilation
FIGURE 9-j (IP). Psychrometric chart — 100 F to 500 F Dry Bulb Temperature
Local Exhaust Ventilation System Design Calculation Procedures
FIGURE 9-k (IP). Psychrometric chart — Up to 1500 F Dry Bulb Temperature
9-65
9-66
Industrial Ventilation
FIGURE 9-l (SI). Psychrometric chart – 0 C to 50 C Dry Bulb Temperature
Local Exhaust Ventilation System Design Calculation Procedures
FIGURE 9-m (SI). Psychrometric chart – 10 C to 120 C Dry Bulb Temperature
9-67
9-68
Industrial Ventilation
FIGURE 9-n (SI). Psychrometric chart – 100 C to 200 C Dry Bulb Temperature
General Industrial Ventilation
10-1
Chapter 10
GENERAL INDUSTRIAL VENTILATION
NOTE: Equations with notation followed by (IP) are designated for inch-pound system only; equations followed by (SI) are designated for
metric use only. If equation bears neither, then it applies to both systems.
10.9.3 Radiation . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10-17
10.1 INTRODUCTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10-3
10.1.1 Generation of Gases and Vapors . . . . . . . . . . . 10-3
10.9.4 Evaporation . . . . . . . . . . . . . . . . . . . . . . . . . . 10-17
10.2 DILUTION VENTILATION PRINCIPLES . . . . . . . . 10-4
10.10 ACCLIMATIZATION OF THE BODY . . . . . . . . . . 10-17
10.3 DILUTION VENTILATION FOR HEALTH . . . . . . . 10-4
10.11 ACUTE HEAT DISORDERS . . . . . . . . . . . . . . . . . . 10-17
10.3.1 General Dilution Ventilation Equation . . . . . . 10-5
10.11.1 Heatstroke . . . . . . . . . . . . . . . . . . . . . . . . . . . 10-17
10.3.2 Mixing Factor (mi) . . . . . . . . . . . . . . . . . . . . . 10-7
10.11.2 Heat Exhaustion . . . . . . . . . . . . . . . . . . . . . . 10-18
10.3.3 Calculating Dilution Ventilation for Steady
State Concentration . . . . . . . . . . . . . . . . . . . . . 10-8
10.11.3 Heat Cramps and Heat Rash . . . . . . . . . . . . . 10-18
10.3.4 Contaminant Concentration Buildup . . . . . . 10-10
10.12 ASSESSMENT OF HEAT STRESS AND
10.3.5 Rate of Purging . . . . . . . . . . . . . . . . . . . . . . . 10-11
HEAT STRAIN . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10-18
10.3.6 Cyclic Operations . . . . . . . . . . . . . . . . . . . . . 10-11
10.12.1 Evaluation of Heat Stress . . . . . . . . . . . . . . . 10-18
10.4 CONFINED SPACE VENTILATION . . . . . . . . . . . . 10-11
10.13
WORKER
PROTECTION . . . . . . . . . . . . . . . . . . . . . 10-19
10.4.1 Reducing Generation Rates . . . . . . . . . . . . . 10-12
10.4.2 Providing Clean Supply Air . . . . . . . . . . . . . 10-12
10.14 VENTILATION CONTROL . . . . . . . . . . . . . . . . . . . 10-19
10.4.3 Blowing into Versus Exhausting from
10.15 VENTILATION SYSTEMS . . . . . . . . . . . . . . . . . . . 10-19
the Space . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10-13
10.16 VELOCITY COOLING . . . . . . . . . . . . . . . . . . . . . . . 10-22
10.5 MIXTURES — DILUTION VENTILATION
10.17 RADIANT HEAT CONTROL . . . . . . . . . . . . . . . . . . 10-22
FOR HEALTH . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10-14
10.18 PROTECTIVE SUITS FOR SHORT EXPOSURES . 10-22
10.6 DILUTION VENTILATION FOR FIRE
AND EXPLOSION (IP UNITS ONLY) . . . . . . . . . . 10-15
10.19 RESPIRATORY HEAT EXCHANGERS . . . . . . . . . 10-22
10.7 FIRE DILUTION VENTILATION FOR MIXTURES 10-16
10.20 REFRIGERATED SUITS . . . . . . . . . . . . . . . . . . . . . 10-23
10.8 VENTILATION FOR HEAT CONTROL . . . . . . . . . 10-16
10.21 ENCLOSURES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10-23
10.9 HEAT BALANCE AND EXCHANGE . . . . . . . . . . . 10-16
10.22 INSULATION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10-23
10.9.1 Conduction . . . . . . . . . . . . . . . . . . . . . . . . . . 10-17
10.9.2 Convection . . . . . . . . . . . . . . . . . . . . . . . . . . 10-17
REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10-23
____________________________________________________________
Figure 10-1
Figure 10-2
Figure 10-3
Effect of Barrier . . . . . . . . . . . . . . . . . . . . . . . .10-6
Effect of Barrier on Poor Conditions . . . . . . . .10-6
Mixing with Uniform Supply and Exhaust
(mi . 1.0) . . . . . . . . . . . . . . . . . . . . . . . . . . . . .10-6
Figure 10-4 Mixing with Uniform Supply and NonUniform Exhaust (2 < mi < 5) . . . . . . . . . . . . .10-7
Figure 10-5 Mixing with Non-Uniform Supply and NonUniform Exhaust on Diagonals (2 < mi < 5) . .10-7
Figure 10-6 Mixing with Exhaust and Supply Only Near
the Floor (4 < mi < 8) . . . . . . . . . . . . . . . . . . . .10-7
Figure 10-7 Mixing with Exhaust and Supply Only Near
the Ceiling (8 < mi < 10) . . . . . . . . . . . . . . . . .10-7
Figure 10-8 Mixing with Supply and Exhaust at Same
End of Room (mi > 10) . . . . . . . . . . . . . . . . . .10-8
Figure 10-9 Mixing Factors Suggested for Inlet and
Exhaust Locations . . . . . . . . . . . . . . . . . . . . . . 10-9
Figure 10-10 Contaminant Concentration Buildup . . . . . . 10-10
Figure 10-11 Rate of Purging . . . . . . . . . . . . . . . . . . . . . . . 10-11
Figure 10-12 Cyclic Generation, Short Cycles . . . . . . . . . . 10-11
Figure 10-13
Figure 10-14
Figure 10-15
Figure 10-16
Figure 10-17
Figure 10-18
Figure 10-19
Figure 10-20
Figure 10-21
Figure 10-22
Figure 10-23
Figure 10-24
Figure 10-25
Cyclic Generation, Long Cycles . . . . . . . . . . 10-12
Avoid Locating Fan Near Opening . . . . . . . . 10-13
Gasoline Powered Fan . . . . . . . . . . . . . . . . . 10-13
Concentrated Exhaust and Poor Mixing . . . . 10-14
Concentrated Supply and Better Mixing . . . 10-14
Blowing Air Into a Confined Space . . . . . . . 10-14
Determination of Wet-Bulb Globe
Temperature . . . . . . . . . . . . . . . . . . . . . . . . . . 10-18
Recommended Heat Stress Alert Limits
(Unacclimatized Workers) . . . . . . . . . . . . . . 10-20
Recommended Heat Stress Exposure Limits
(Acclimatized Workers) . . . . . . . . . . . . . . . . 10-20
Good Natural Ventilation and Circulation . . 10-21
Good Mechanically Supplied Ventilation . . . 10-21
Spot Cooling with Volume and Directional
Control . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10-22
Heat Shielding . . . . . . . . . . . . . . . . . . . . . . . . 10-23
10-2
Table 10-1
Table 10-2
Industrial Ventilation
Effective Volumetric Flow Rates (Q')
for Vapors per Pint of Evaporated Liquid . . . 10-5
Estimating Energy Consumed by
Task/Work Performed . . . . . . . . . . . . . . . . . . 10-19
Table 10-3
Table 10-4
Acceptable Comfort Air Motion at the
Worker . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10-22
Relative Efficiencies of Common
Shielding Materials . . . . . . . . . . . . . . . . . . . . 10-22
General Industrial Ventilation
10.1
INTRODUCTION
General industrial ventilation is a broad term that refers to
the exhaust and supply of air with respect to an area, room, or
building. It can be divided further into specific functions as
follows:
1) Dilution Ventilation is the dilution of contaminated air
with uncontaminated air for the purpose of controlling
potential airborne health hazards, fire and explosive
conditions, odors, and other contaminants. Dilution
ventilation can also include the control of airborne contaminants (vapors, gases, and particulates) generated
within tightly constructed buildings.
Dilution ventilation is not as effective for health
hazard control as is local exhaust ventilation.
Circumstances may be found in which dilution ventilation system provides an adequate amount of control
more economically than a local exhaust system. Do not
base the economic considerations entirely upon the
first cost of the system since dilution ventilation frequently exhausts large amounts of heat from a building,
and may increase the energy cost of the operation.(10.1)
2) Heat Control Ventilation is the control of indoor atmospheric conditions associated with hot industrial environments such as are found in foundries, laundries,
bakeries, etc., for the purpose of preventing acute discomfort or injury.
10.1.1 Generation of Gases and Vapors. Contaminant
gases and vapors may be released from storage cylinders
deliberately or inadvertently as part of a process or they may
be created by chemical reactions in a process.
Because vapors are due to evaporation of a liquid, their generation rate increases directly with surface area exposed to the
air and temperature.
The vapor concentration in the space above a liquid in a
closed container in which liquid remains is equal to the vapor
pressure at that temperature divided by the pressure of all the
gases and vapors in that space. If the container space contains
a mixture of air and the vapor is at atmospheric pressure, then
the concentration of the vapor will be the vapor pressure divided by the atmospheric pressure. For example, if the vapor pressure at 70 F [21 C] is 203.5 "wg [380 mmHg], then at an atmospheric pressure of 407 "wg [760 mmHg], the saturation concentration of the vapor would be 203.5/407 [380/760] (i.e.,
50% or 500,000 ppm).
If the air/vapor mixture is withdrawn from that space and
replaced by uncontaminated air, the space concentration will
fall below the saturation concentration. The concentration will
gradually rise back to saturation concentration as long as some
liquid remains in the container. That space concentration will
exponentially approach saturation concentration with time.
That is, the change in concentration with time will become
slower and slower as the concentration increases.
10-3
Evaporation rate is directly related to concentration in the
air relative to the saturation concentration. At any time, some
molecules are condensed to liquid state while others are escaping the surface by evaporation. At lower concentrations, the
fraction evaporating is much greater than the fraction condensing and the concentration increases rapidly. At higher concentrations, the two rates become more and more equal and the
concentration increases correspondingly slower. At equilibrium, the two processes are equal and the concentration does not
change at all or bounces back and forth within a narrow range.
If the temperature changes, the vapor pressure changes
markedly. The vapor pressure can never quite reach zero
(frozen into a solid) or atmospheric pressure (100% concentration in the space above the liquid). As the vapor pressure
changes with temperature, the saturation concentration changes
proportionally. The space concentration in a closed container
can be predicted very accurately from its temperature.
The time required to re-establish the saturation concentration will fall sharply if the liquid surface is agitated.
Outside of a closed container (e.g., a puddle of liquid), there
is a thin layer of air that adheres to any surface, thus limiting
the exchange to the room air. The concentration in that layer
will generally be much higher than the air even a few millimeters above it.
However, room air currents flowing over the puddle
increases the exchange rate with the thin layer, greatly increasing the generation rate. That is why fanning a puddle of liquids
with moderate vapor pressures (e.g., ethanol) will greatly
increase their rate of evaporation (or cooling). Hence, blowing
dilution air over a liquid surface will greatly increase evaporation rate.
Since evaporation occurs from surfaces, generation (or
evaporation, or sublimation) rate due to evaporation is directly
proportional to the total surface area exposed to room air. In
this regard, it is vital to understand that agitating a surface
increases its surface area (and breaks up the thin layer of air).
If the agitation is violent enough to produce particles of liquid,
the surface area is vastly increased and the generation rate
increases dramatically.
For a mixture of evaporating liquids, the rate of evaporation
of each one is affected by the other components. The concentration in the air of each will be governed by its vapor pressure.
If the liquid evaporates completely, then the concentration of
each is proportional to the number of moles created from the
mass of each as it is evaporated.
The amount of vapor generation per surface area of liquid is
difficult to predict even for quiet, undisturbed surfaces. For a
surface disturbed by room air currents or agitation, it can be
determined for an exact set of conditions empirically but cannot be modeled so that that knowledge can be extrapolated to
predict generation rates for a different set of conditions.
The change in air concentration of a contamination can be
10-4
Industrial Ventilation
dramatic if there is a large spill. In minutes the concentration
can jump from barely measurable to deadly if the chemical is
both volatile and highly toxic. Even if it is not notably volatile
or toxic, the chemical can evaporate to dangerous air concentrations if it is spread over a large surface area or if it is
released in a relatively small confined space.
As discussed earlier, air drawn across the surface of an evaporating liquid strips away the saturated layer of vapor, accelerating the evaporation rate (G). Thus, it is difficult to estimate G
when general ventilation systems are drawing air over a spill or
through a confined space that still contains the evaporating liquid because the evaporation rate itself will be a function of the
ventilation rate. The designer must estimate the time required
for complete evaporation under the prevailing conditions. Air
volume (Q) should be estimated based on the assumption that
the liquid will evaporate quickly and should estimate timeuntil-safe entry on the assumption that it evaporates slowly.
10.2
DILUTION VENTILATION PRINCIPLES
The principles of dilution ventilation system design are as
follows:
1) Select from available data the amount of air required for
satisfactory dilution of the contaminant. The values tabulated in Table 10-1 assume perfect distribution and
dilution of the air and solvent vapors. These values must
be multiplied by the selected mixing factor (mi) (see
Section 10.3.2).
2) Locate the exhaust openings near the sources of contamination, if possible, in order to obtain the benefit of
local ventilation.
3) Locate the air supply and exhaust outlets so that the air
passes through the zone of contamination. The operator
should remain between the air supply and the source of
the contaminant.
4) Replace exhausted air by use of a powered supply air
system. This supply or replacement air should be heated
or possibly cooled to satisfy the comfort requirements of
the space. Dilution ventilation systems usually handle
large flows of air by means of low pressure fans (Figures
10-14 and 10-15). Adequate quantities of supply air must
be provided if the system is to operate satisfactorily.
5) Avoid re-entry of the exhausted air by discharging the
exhaust high above the roof line or by assuring that no
window, outdoor supply air intakes, or other such
openings are located near the exhaust discharge.
6) Minimize any opportunity for flexible duct to kink or
become twisted. Use rigid duct where possible (can
include rigid spiral duct).
10.3
DILUTION VENTILATION FOR HEALTH
In general, the concentration of a contaminant is not uni-
form in a room with contaminant sources. It will vary with
location within a room because:
1) Sources are not spread uniformly over the contaminated area. Some locations of interest are much closer to
sources than others.
2) Some sources are mobile.
3) The supply air is usually not distributed uniformly in
the ventilated area.
4) The exhaust air is not removed uniformly from the ventilated area.
5) There are competing air motions due to crossdrafts
traffic, machinery or human motion that can move contaminants a large range of distances.
Thus, concentration will vary with proximity to
sources and with the direction of air movement near
sources. The effect of distance from the source on concentrations is predictable only under highly controlled
conditions seldom found in the workplace. In the
absence of air movement, the contaminant concentrations will be equal in all directions at a given distance
from the source. Under these idealized conditions the
contaminant spreads solely by diffusion, producing a
distribution of decreasing concentrations with increasing distances from the source.
The use of dilution ventilation for health hazards has four
limiting factors: 1) the quantity of contaminant generated must
not be too great or the airflow rate necessary for dilution will
be impractical; 2) workers must be far enough away from the
contaminant source or the generation of contaminant must be
in sufficiently low concentrations so that workers will not have
an exposure in excess of the established Threshold Limit
Values (TLVs® should be used as guidelines only and not as
absolute criteria for a safe and acceptable workplace); 3) the
toxicity of the contaminant must be low (substances of
unknown toxicity should be treated as highly toxic until
proven otherwise); 4) the generation of contaminants must be
reasonably uniform.
Dilution ventilation is used most often to control the vapors
from organic liquids with a TLV® of 20 ppm or higher. In order
to successfully apply the principles of dilution to such a problem, factual data are needed on the rate of vapor generation or
on the rate of liquid evaporation. Usually such data can be
obtained from the plant if adequate records on material consumption are kept.
The general movement of air may either increase or reduce
the exposures to workers near sources of contamination. In
workplaces, air is generally distributed by convection currents
and the effects of nearby moving objects. If the disturbances
are symmetrical and vigorous, they may cause the contaminated air to be mixed thoroughly and there are only moderate differences in exposures at different locations. In that condition,
the relative location of the worker and the source could be
General Industrial Ventilation
10-5
TABLE 10-1. Effective Volumetric Flow Rate (Q') for Vapors per Pint of Evaporated Liquid
unimportant. It is also possible that there is a discernible crossdraft pushing the contaminant in one direction. In that case, the
concentrations would be highest in the direction the crossdraft
pushes the contaminant. As the crossdraft moves along it will
tend to mix with surrounding, cleaner air. For that reason, the
concentrations may fall with increasing distance from the
source in the direction of the air movement. The effect of
directionality on worker exposures depends on the relative
position of the workers and the direction of the contaminant
movement as follows:
1) If the contaminant cloud moves towards the worker,
the sweeping movement of air would tend to increase
exposures.
2) If clean air moves towards a worker who is very near
the contaminant source, it may sweep the contaminant
away from the worker.
3) If the clean air moves towards the worker’s back when
the contaminant is in front of the worker, a wake zone
downstream of the worker’s body may circulate the air
towards the worker’s face, sometimes producing high
exposures.
It is sometimes useful to take advantage of air sweeping to
reduce exposures by blowing clean air towards an exposed
worker, but this strategy can fail if the direction of air movement is in the wrong direction.
It is also likely that the sweeping strategy will fail if the
worker is within a couple of arm’s lengths from the source. As
is discussed later, air blowing towards a worker’s back will
create wake zones that can draw contaminants back to the
worker, sometimes substantially increasing exposures.
The most prudent strategy is to locate the source extremely
close to the exhaust point so that very little contaminated air
escapes to the general area of the room. In this strategy, the
exhaust point would act as a capturing hood (see Chapter 6)
and dilution issues would not be particularly relevant. It is difficult to produce effective control using the sweeping strategy
unless the source is very close to an exhaust point. In particular, if the source is well away from an exhaust point, air movement is likely to vary in direction and magnitude in ways that
are very difficult to control. It is difficult due to the variable
flow conditions in workplaces. Even when the air in the room
apparently flows from left to right, and presumably sweeps the
contaminated air from left to right with it, the concentration to
the left of the source is NOT zero because of the eddy currents
induced by the entrainment. Turbulent flow of air across large
spaces tends to induce large eddy currents that can move contaminants counter to the sweep direction.
Eddy currents can also be produced when the supply air
moves around obstructions in its path (Figures 10-1 and 10-2)
to the fan and by movement of process machinery, conveyors,
tow motors and the body motions of the operators. For example, a pedestrian walking near a source at normal speeds
(150–250 fpm) [0.76–1.3 m/s] will create a moving wake
behind him that can easily draw the contaminated air from the
source towards himself.
10.3.1 General Dilution Ventilation Equation. The ventilation rate needed to maintain a constant concentration at a uniform generation rate is derived by starting with a fundamental
material balance (assuming no contaminant in the air supply).
Rate of Accumulation = Rate of Generation –
Rate of Removal
or
10-6
Industrial Ventilation
Q
QN = __
mi
[10.3]
where: Q = actual ventilation rate, acfm [am3/s]
QN = effective ventilation rate, acfm [am3/s]
mi = a factor to allow for incomplete mixing
(Figures 10-3 through 10-9)
Equation 10.2 then becomes:
Q = (mi)(G/Cg)(106)
This mixing factor (mi) is based on several considerations:
1) The distribution of supply air introduced into the room
or space being ventilated (Figure 10-9) and how well it
mixes with room air.
FIGURE 10-1. Effect of barrier
Vr dCg = G dt – QNCg dt
[10.1]
where: Vr = volume of room, ft3 [m3]
G = rate of generation of contaminant, lbm/min [g/s]
QN = effective volumetric flow rate, acfm [am3/s]
Cg = concentration of gas or vapor at time t,
ppm [ppm]
t = time
At a steady state,
dCg = 0
G dt = QNCg dt
At a constant concentration (Cg) and uniform generation rate
(G), then
G(t2 – t1) = QNCg (t2 – t1)
QN = (G/Cg)(106)
[10.2]
2) The toxicity of the solvent. Although TLVs® are only
guidelines for toxicity levels and TLVs® and toxicity
are not synonymous, the following guidelines have
been suggested for increasing the appropriate mi value:
Slightly toxic material: TLV® > 500 ppm
Toxic material: TLV® # 20–500 ppm
Highly toxic material: TLV® < 20 ppm
3) The judgment of any other circumstances that an industrial hygienist determines to be important based on
experience and the specific workspace. Included in
these criteria are such considerations as:
a) Duration of the process, operational cycle, and normal locations of workers relative to sources of contamination.
b) Location and number of points of generation of the
contaminant in the workroom or area.
c) Seasonal changes in the amount of natural
ventilation.
d) Operational effectiveness of the mechanical air
moving devices.
Due to incomplete mixing, a mixing factor (mi) is introduced
to the rate of ventilation:
FIGURE 10-2. Effect of barrier on poor conditions
[10.4]
FIGURE 10-3. Mixing with uniform supply and exhaust
(mi . 1.0)
General Industrial Ventilation
FIGURE 10-4. Mixing with uniform supply and non-uniform
exhaust (2 < mi < 5)
e) Other circumstances that may affect the concentration of hazardous material in the breathing zone of
the workers.
The mixing factor selected, depending on the above considerations, ranges from 1 to 10.
10.3.2 Mixing Factor (mi). The concentrations in a room
with a contaminant source will vary substantially with location.
The mixing factor is defined as a de-rating of a distribution system. This was named “K” in previous editions of the Manual
and is known to some as the K factor. Not unlike the System
Effects encountered with fans, it is a technique to de-rate the
effectiveness of the mixing and uses unity as the best mixing
condition in a space. With mi equal to 1.0, the amount of ventilation required would equal the values obtained when there is
perfect mixing and supply air properly reduces the exposure
limits to desired values. Values greater than 1.0 reflect imperfect mixing of the supply air and removal of pollutants and thus
higher required air supply and exhaust volumes.
10-7
FIGURE 10-6. Mixing with exhaust and supply only near
the floor (4 < mi < 8)
Even with perfect mixing (mi = 1) it is likely that different
workers would still have different exposures at the same location due to differences in physiology, vigor of movements, and
work practices. Values of mi are difficult to predict without
computational fluid dynamics (CFD) simulations (see Chapter
12, Section 12.2). However, it is not difficult to rank order values of mi within different layouts. As a general rule, having a
source and a worker both within an area of poor mixing leads
to high exposures to that worker and thus to high values of mi
at that location. If the worker is handling the source, the value
of mi should be assumed to be higher. Typical values for mixing factors are shown in Figure 10-9. Note that these are estimates only and not based on any laboratory or other research
values.
The ideal design is a long, narrow tunnel with a uniform
exhaust and uniform supply air at each end (Figure 10-3). The
concentrations in the space would probably not vary appreciably due to the mixing, so if the worker is not near the source
or is well upstream of it, values should be close to 1.0.
For specific locations within the room, the closer mi
approaches unity the more thorough the mixing. Values of mi
> 1.0 indicate imperfect mixing. For poor mixing, mi would be
higher than 1 (e.g., 2–10). Please note that systems handling
potentially toxic materials should employ mixing factors at the
higher end of ranges.
Having uniformity of flow at only one end of the room does
not ensure that flow is uniform elsewhere in the space (Figure
10-4). Based on location of the worker, values can approach
1.0, but if located down near the corner a higher value mi
(5–10) might be required.
FIGURE 10-5. Mixing with non-uniform supply and non-uniform exhaust on diagonals (2 < mi < 5)
FIGURE 10-7. Mixing with exhaust and supply only near
the ceiling (8 < mi < 10)
10-8
Industrial Ventilation
tration (Cg) expressed in parts per million (ppm) is considered
to be the Threshold Limit Value (TLV®). For liquid solvents,
the rate of generation is
CONSTANT H SG H ER
G = ___________________
MW
where:
G = generation rate, acfm [am3/s]
CONSTANT = 403 ft3-lbm/pt-lbmol [24 m3-g/l-gmol]
SG = specific gravity of volatile liquid
ER = evaporation rate of liquid, pts/min [l/s]
FIGURE 10-8. Mixing with supply and exhaust at same end
of room (mi > 10)
MW = molecular weight of liquid, lbm/lbmol
[g/gmol]
Cg = acceptable concentration, ppm [ppm]
Diagonally placed supply and exhaust points produces still
larger stagnant sections of the room that are poorly mixed
(Figure 10-5). Mixing factors in this case again are dependent
on the location of the worker. Near supply source and exhaust
point, the values are approximately 2–3. Other worker locations may require values as high as 6 to 10.
Having the single point supply and exhaust both towards
the bottom produces a large stagnant section across the top of
the room (Figure 10-6). Mixing factors would be estimated in
the range of 4 to 8.
Similarly placing exhaust and supply at high locations at
opposite ends (Figure 10-7) is likely to produce large stagnation areas near the floor, affecting workers much more (mixing
factors between 8 and 10).
When the single exhaust and single supply are on the same
end (Figure 10-8), the other end of the room is likely to be
highly stagnant. If the source is on the ventilated end, the
results may be acceptable. If the source and a worker are located on the stagnant opposite end, the worker’s exposure is likely to be high and mixing factor could be 9 or 10.
Barriers that block the flow can either improve or degrade
the mixing in critical areas, depending on whether the barriers
channels flow away from the source or through it. In Figure
10-2 a worker next to a source on the right side of the barrier
could be heavily exposed and mi = 10. On the left side of the
barrier mi could be as low as 1 to 2.
10.3.3 Calculating Dilution Ventilation for Steady State
Concentration. The concentration of a gas or vapor at a steady
state can be expressed by the material balance equation shown
in Equation 10.2:
QN = ( G / Cg)(106)
Therefore, the rate of flow of uncontaminated air required to
maintain the atmospheric concentration of a hazardous material at an acceptable level can be easily calculated if the generation rate can be determined. Usually, the acceptable concen-
Thus, QN = (G/Cg)(106) can be expressed as:
403 H 106 H SG H ER
QN = _________________
MW H Cg
[10.5] IP
24 H 106 H SG H ER
[QN = _________________ ]
MW H Cg
[10.5] SI
EXAMPLE PROBLEM 10-1 (Dilution Airflow with
Constant Evaporation of Contaminant)
Methyl chloroform is lost by evaporation from a tank at a rate
of 1.5 pints [0.71 l] per 60 minutes [3600 s]. What is the effective ventilation rate (QN) and the actual ventilation rate (Q)
required to maintain the vapor concentration at the TLV®?
TLV = 350 ppm, SG = 1.32, MW = 133.4, Assume mi = 5
Assuming perfect dilution, the effective ventilation rate (QN) is
QN =
(403)
(106) (1.32) (1.5/60)
____________________
= 285 acfm
(133.4) (350)
6
24 (10 ) (1.32) (0.71/3600)
[QN = ______________________ = 0.13 am3/s]
(133.4) (350)
Due to incomplete mixing (mi = 5) the actual ventilation rate (Q) is
(403) (106) (1.32) (1.5/60) (5)
Q = _______________________ = 1,425 acfm
(133.4) (350)
6
24 (10 ) (1.32) (0.71/3600) (5)
[Q = _________________________ = 0.67 am3/s]
(133.4) (350)
General Industrial Ventilation
10-9
10-10
Industrial Ventilation
10.3.4 Contaminant Concentration Buildup (Figure 10-10).
The concentration of a contaminant can be calculated after
any time interval. Rearranging the differential material
balance (Equation 10.1) results in
dCg
dt
_______
= __
Vr
G – QNCg
which can be integrated to yield
[10.6]
where subscript 1 refers to the initial condition and subscript 2
refers to the final condition. If it is desired to calculate the time
required to reach a given concentration, rearranging gives t2 –
t1, or Dt.
FIGURE 10-10. Contaminant concentration buildup
[10.7]
If Cg = 0, then the equation becomes
1
[10.8]
NOTE: The concentration Cg2 is ppm or parts of
contaminant/10 6 parts of air (e.g., if Cg2 is in units of ppm,
enter Cg2 as 200/106).
If it is desired to determine the concentration level (Cg2)
after a certain time interval, t2 – t1 or Dt, and if Cg1 = 0, then
the equation becomes
[10.9]
[SI units] Methyl chloroform is being generated under the following conditions: G = 0.0005663 am3/s, Q = 2.831 am3/s, Vr
= 2832 m3, Cg1 = 0, Cg2 = 200 ppm or 200)106, mi = 3.
6
NOTE: To convert Cg2 to ppm, multiply the answer by 10 .
EXAMPLE PROBLEM 10-2 (Time to Reach a Concentration with Constant Evaporation of Contaminant)
[IP units] Methyl chloroform is being generated under the
following conditions: G = 1.2 acfm, Q = 6,000 acfm, Vr =
100,000 ft3, Cg1 = 0, Cg2 = 200 ppm or 200)106, mi = 3.
Using the same value, what will the concentration be after 60
minutes?
Using the same value, what will the concentration be after 60
minutes [3,600 seconds]?
General Industrial Ventilation
10-11
10.3.5 Rate of Purging (Figure 10-11). Where a quantity of
air is contaminated but where further contamination or generation has ceased, the rate of decrease of concentration over a
period of time is as follows:
Vr dCg = – QNCg dt
or,
[10.10]
FIGURE 10-11. Rate of purging
EXAMPLE PROBLEM 10-3 (Dilution of Contaminant
Concentration after Removal of Source)
In the room of the example in Section 10.3.3, assume that
ventilation continues at the same rate (QN = 2,000 acfm), but
that the contaminating process is interrupted. How much time
is required to reduce the concentration from 100 (Cg1) to 25
(Cg2) ppm?
In the problem above, if the concentration (Cg1) at t1 is 100
ppm, what will concentration (Cg1) be after 60 minutes (Dt)?
10.3.6 Cyclic Operations. Contaminant generation is usually cyclic, especially in batch operations. If the cycles are short
compared to the total buildup time, the concentrations may be
only marginally higher and lower than they would have been if
the same amount of contaminant were released at a constant rate
(Figure 10-12). However, if the cycles are very long compared
to the buildup period, then the peaks and valleys diverge substantially from the continuous exponential curve. In investigating the 8-hour time-weighted average (TWA) exposure, these
deviations are of little concern because the average will be the
same as long as the buildup period is less than eight hours.
For short-term averages, cycles can become very important,
especially in the event that the cycles are longer than the measurement period. For example, if the peak-to-peak time in
Figure 10-13 is one hour long, then a 15-minute TWA taken at
the peak could be much higher than the 15-minute TWA taken
at the trough.
10.4
CONFINED SPACE VENTILATION
When considering whether a space should be considered a
confined space for purposes of ventilation design, a functional
definition is useful. A confined space is a location: 1) where concentrations of hazardous air contaminants can build up rapidly
due to its relatively small size and limited air exchange with the
outside, and 2) a person will enter the space. Using this definition, a 10 ft × 10 ft utility room housing chlorine cylinders is a
confined space if the cylinders are leaking or there is a signifi-
FIGURE 10-12. Cyclic generation, short cycles
10-12
Industrial Ventilation
(G) must be greater than zero, which means that contaminant
is still evolving within the confined space. If the concentration
is unacceptable, increasing the airflow to a higher level should
reduce the concentration linearly:
[10.13]
FIGURE 10-13. Cyclic generation, long cycles
cant risk they could leak. Likewise, a 7 ft deep trench in which
workers are standing at the bottom and spraying epoxy could be
considered a confined space. The space is even more hazardous
if egress is difficult. Mixing tanks with entry ports just large
enough for a worker to enter and exit via a ladder are classic
examples of the latter.
Monitoring, purging, and ventilation are the keys to protecting
workers from airborne contaminants in confined spaces.
However, ventilation is not sufficient by itself to keep confined
space entries safe. The OSHA Confined Space Standards (29
CFR 1910.146 and 1926 Subpart AA) are a basis for a standard
operating procedure and requires a permit. Ensure that the latest
issue of the standard and exposure information is being used.
Although this is a ventilation text, monitoring is discussed since
it is affected by ventilation practices.
An important value to compute is the time required to
achieve seven air changes (t7AC)
[10.11]
The time computed using Equation 10.11 is a useful benchmark and is a reasonable estimate of a minimum time for safe
entry for most situations if no contaminant is being generated
and the supply air is clean.
If concentrations reach a significant non-zero concentration
in the exhaust air and do not fall lower for a substantial length
of time, then steady-state conditions have been achieved. The
time-dependent terms in Equation 10.1 become zero and this
is expressed by:
Cexh = [(G/Q')(106) + Csup]
[10.12]
where:
Cexh = concentration of exhaust gas (ppm) [ppm]
G = generation rate (acfm) [am3/s]
Q' = effective volumetric flow rate (acfm) [am3/s]
Csup = vapor concentration in supply air (ppm) [ppm]
If monitoring confirms that Csup is zero, then generation rate
where:
Cexh = original concentration of exhaust gas (ppm)
[ppm]
Q = exhaust flow rate (acfm) [am3/s]
Ct = target concentration of exhaust gas (ppm)
[ppm]
10.4.1 Reducing Generation Rates. If an existing system
inadequately protects workers it is possible to reduce the exposures to acceptable levels by:
1) Draining liquids completely.
2) Minimizing the potential for spills and leaks from the
work.
3) Installing local exhaust hoods within the space to collect contaminants released by the work (see Chapter 6).
4) Locating the exhaust and supply air intake and discharge points (see sections below) so that airflow patterns are effective in mixing the contaminated air with
the ducted air (see Section 10.4.3). If the initial concentration is very high or if the room containing the confined space is relatively small or poorly ventilated it
may be advisable to start with negative pressure ventilation and switch to positive pressure.
5) If experience or logic indicates that some sections of
the space will have concentrations higher than the average for the total volume, move the flexible duct to that
location for some fraction of the initial purge period.
6) Locating clean air intake points where there are no contaminants.
7) Purging the space with at least the minimum level of
airflow and for at least the minimum time period determined from relevant experience. Do not allow entry
until after the minimum time has passed and measured
concentrations are below target levels. Both requirements should be met (not just one).
8) Continue ventilating the space with at least the minimum
level of airflow determined from relevant experience.
9) If the concentration is still unacceptably high after an
unacceptably long period of time increase the airflow
volume (Q) to the point that the highest concentration
within the space is lower than the target concentration
(Ct).
10.4.2 Providing Clean Supply Air. The air blown or
drawn into a confined space must be free of hazardous contaminants. In general:
General Industrial Ventilation
10-13
1) Locate fresh air intakes for blowers well away from
potential sources of contamination, including the air
pouring out of the confined space if air is blown into
the space (Figure 10-18).
2) Remove sources of exposures from the vessel’s access
hole if air will be drawn into that opening. If that cannot be done, blow air into the vessel and ensure that the
fan intake draws clean air.
When blowing air into a space, it is necessary only to place
the fan in an area having non-contaminated air (clean area) or
connect inlet ductwork to it and place its open end to a clean
area (Figure 10-18). If the fan is placed on or near the vessel,
it is important to run inlet ductwork as far as necessary to reach
a clean area.
When exhausting air from a space, the inlet port of the space
becomes the path of clean air. If the fan is placed near the port
(Figure 10-14), outlet ductwork should be employed to carry
the contaminated air to a space that is either uninhabited or is
large enough and has enough ventilation that excessive concentrations will not build up.
In cases where a gasoline-powered fan is employed (Figure
10-15), the effluent from the engine exhaust is a serious hazard. The engine exhaust should not be near the air intake of the
blower for positive pressure systems or near the vessel inlet
port for negative pressure systems.
The range of concentrations within a ventilated space is
strongly influenced by the mixing factor (mi) (see Section
10.3.2). It is difficult to predict the mixing factor for a volume
without extensive sampling data taken over time. Mixing factors normally apply to pinpoint locations within the space and
not to the entire space.
The range of concentrations within a space can be very
large. This is especially true for the typical case where the contaminant is not spread uniformly throughout the space. The
FIGURE 10-15. Gasoline powered fan
less diffuse the source, the greater the concentration gradients
within the space. Likewise, if the worker is very close to the
source, the mixing factor will be high.
Mixing factors (mi) typically range from low (< 2.0) to high
(> 5) within the same space. A lower range is sometimes
achieved by placement of the exhaust points and supply points
in relationship to the sources. Blowing air into a space at high
velocities will produce better mixing because of attendant high
kinetic energies. High velocities can produce worker complaints especially if the supply air is relatively cool. Even in
areas with strong sources, high velocity supply air can blow
the contaminant into adjacent clean areas.
10.4.3 Blowing into Versus Exhausting from the Space.
In most cases, airflow should be blown into a confined space
rather than exhausting from it. This is generally advisable
because air drawn into the confined space is likely not to mix
well with the contaminated air in the space. This will increase
the mixing factor (mi) and the required flow rate (Q) and/or
purge time. It also is important to blow air into the confined
space (vessel, tank, or enclosure) if a significant source is near
the opening to the confined space (e.g., exhaust pipe from a
vehicle).
As shown in Figure 10-16, the air exhausted from the room
has little kinetic energy to stir the surrounding air. Note that
even the pressure system can still have poor mixing in parts of
the space (Figure 10-17).
Three important exceptions where exhaust design shown in
Figure 10-18 is preferred are:
1) During initial purging of highly contaminated spaces if
a worker must be outside the confined space near the
port where the air would be exhausted.
FIGURE 10-14. Avoid locating fan near opening
2) During initial purging of highly contaminated spaces if
the room with the vessel is small enough and poorly
10-14
Industrial Ventilation
FIGURE 10-16. Concentrated exhaust and poor mixing
ventilated enough to become significantly contaminated by the vessel effluent.
3) If a tent or other shelter is installed above the vessel’s
exhaust opening. The air concentration inside the shelter will be about the same as the air discharging from
the confined space if positive pressure ventilation is
employed.
FIGURE 10-18. Blowing air into a confined space
Air should be exhausted from the space until the concentration levels are low and safe enough to discharge to the room.
The exhaust air should be released into a room that is either
unoccupied or is large enough or well ventilated enough to
prevent excessive concentration buildup. If those conditions
cannot be met then use an exhaust system to clear the area.
10.5
MIXTURES — DILUTION VENTILATION FOR
HEALTH
The parent liquid for which dilution ventilation rates are
being designed might consist of a mixture of solvents.
When two or more hazardous substances have similar toxicological effect on the same target organ or system, their combined
effect, rather than that of either individually, should be given primary consideration. In the absence of information to the contrary, the effects of the different hazards should be considered as
additive where the health effect and target organ or system is the
same. That is, if the sum of the following fractions,
FIGURE 10-17. Concentrated supply and better mixing
exceeds unity, then the threshold limit of the mixture should
be considered as being exceeded, where Cg indicates the
observed atmospheric concentration and TLV® indicates the
corresponding threshold limit value.(10.3)
In the absence of information to the contrary, the dilution
ventilation should, therefore, be calculated on the basis that the
effect of the different hazards is additive. The air quantity
required to dilute each component of the mixture to the
required safe concentration is calculated, and the sum of the air
quantities is used as the required dilution ventilation for the
mixture.
Exceptions to the above rule may be made when there is
reason to believe that the chief effects of the different harmful
substances are not additive but independent, as when purely
local effects on different organs of the body are produced by
the various components of the mixture. In such cases, the
threshold limit ordinarily is exceeded only when at least one
member of the series itself has a value exceeding unity, e.g.,
Therefore, where two or more hazardous substances are present and it is known that the effects of the different substances
are not additive but act independently on the different organs
of the body, the required dilution ventilation for each component of the mixture should be calculated and the highest acfm
[am3/s] obtained should be used as the dilution ventilation rate.
General Industrial Ventilation
EXAMPLE PROBLEM 10-4 (Dilution Airflow with
Constant Evaporation of Two Contaminants)
A cleaning and gluing operation is being performed; methyl
ethyl ketone (MEK) and toluene are both being released. Two
pints [0.946] of each are being released every 60 min. Select
an mi value of 4 for MEK and an mi value of 5 for toluene; SG
for MEK = 0.805, for toluene = 0.866; MW for MEK = 72.1,
for toluene = 92.13. Both have narcotic properties, and the
effects are considered additive. Air samples disclose concentrations of 150 ppm MEK and 50 ppm toluene. The sum of
(Cg1/TLV1) and (Cg2/TLV2) exceeds unity (1); therefore, the
TLV® of the mixture is exceeded. The volumetric flow rate at
standard conditions required for dilution of the mixture to the
TLV® would be as follows:
(403) (0.805) (106) (4) (2/60)
Q for MEK = ______________________ = 3,000 acfm
72.1 H 200
(24) (0.805) (106) (4) (0.946/3600)
[Q for MEK = __________________________ = 1.41 am3/s]
72.1 H 200
(403) (0.866) (106) (5) (2/60)
Q for toluene = ______________________ = 12,627 acfm
92.13 H 50
(24) (0.866) (106) (5) (0.946/3600)
[Q for toluene = ___________________________
93.13 H 50
= 5.86 am3/s]
10-15
Equation 10.5 [IP] can be modified to yield air quantities to
dilute below the LEL. By substituting LEL for TLV®:
(403) (SG) (100) (ER) (Sf)
Q = _____________________ (for Standard Air)
(MW)(LEL)(B)
[10.14] IP
NOTES: 1. Since LEL is expressed in percent (parts per 100)
rather than ppm (parts per million as for the
TLV®), the coefficient of 1,000,000 becomes 100.
2. Sf is a safety coefficient that depends on the percentage of the LEL necessary for safe conditions.
In most ovens and drying enclosures, it has been
found desirable to maintain vapor concentrations
at not more than 25% of the LEL at all times in all
parts of the oven. In properly ventilated continuous ovens, a minimum Sf coefficient of 4 (25% of
the LEL) is used. In batch ovens, with good air
distribution, the existence of peak drying rates
requires an Sf coefficient of a minimum of 10 or
12 to maintain safe concentrations at all times. In
non-recirculating or improperly ventilated batch
or continuous ovens, even larger Sf coefficients
may be necessary.
3. B is a constant that takes into account the fact that
the lower explosive limit of a solvent vapor or air
mixture decreases at elevated temperatures. B = 1
for temperatures up to 250 F; B = 0.7 for temperatures above 250 F.
Q for mixture = 3,000 + 12,627 = 15,627 acfm
[Q for mixture = 1.41 + 5.86 = 7.27 am3/s]
EXAMPLE PROBLEM 10-5 (Dilution Airflow to Avoid
Explosive Mixture with Constant Evaporation of
Solvent) (IP Units Only)
10.6
DILUTION VENTILATION FOR FIRE AND
EXPLOSION (IP UNITS ONLY)
Another function of dilution ventilation is to reduce the concentration of vapors within an enclosure to below the lower
explosive limit (LEL). It should be stressed that this concept is
never applied in cases where workers are exposed to the vapor.
In such instances, dilution rates for health hazard control are
always applied.
The TLV® of xylene is 100 ppm. The LEL of xylene is a 1%
content ratio or 10,000 ppm. An atmosphere of xylene safeguarded against fire and explosion usually will be kept below
25% of the LEL, or 2,500 ppm. Exposure to such an atmosphere may cause severe illness or death. However, in baking
and drying ovens, in enclosed air drying spaces, within ventilation ducts, etc., dilution ventilation for fire and explosion is
used to keep the vapor concentration to below the LEL.
A batch of enamel dipped shelves is baked in a recirculating
oven at 350 F for 60 minutes. Volatiles in the enamel applied to
the shelves consist of two pints of xylene. What oven ventilation rate, in acfm, is required to dilute the xylene vapor concentration within the oven to a safe limit at all times?
For xylene, the LEL = 1.0%; SG = 0.88; MW = 106; Sf = 10; B
= 0.7.
From Equation 10.14 [IP]:
(403) (0.88) (100) (2/60) (10)
Qstd = _______________________ = 159 scfm
(106)(1.0)(0.7)
Since the above equation is at standard conditions, the airflow rate must be converted from 70 F to 350 F (operating
conditions):
Qstd = (scfm) (df)
10-16
Industrial Ventilation
(460 F + 350 F)
= (acfmstp) _____________
(460 F + 70 F)
= 243 acfm @ 350 F
EXAMPLE PROBLEM 10-6 (Dilution Airflow to Avoid
Explosive Mixture with Varying Evaporation of Solvent)
(IP Units Only)
In many circumstances, solvent evaporation rate is nonuniform due to the process temperature or the manner of
solvent use.
A 6 ft diameter muller is used for mixing resin sand on a 10minute cycle. Each batch consists of 400 pounds of sand, 19
pounds of resin, and 8 pints of ethyl alcohol (the ethyl alcohol
evaporates in the first two minutes). What ventilation rate is
required?
For ethyl alcohol, LEL = 3.28%; SG = 0.789; MW = 46.07;
Sf = 4; B = 1.
(403) (0.789) (100) (8/2) (4)
Qstd = ______________________ = 3,367 scfm
(46.07)(3.28)(1)
Another source of data is the NFPA (National Fire
Protection Association) 86, Standard for Ovens and
Furnaces.(10.4) This contains a more complete list of solvents and their properties. In addition, it lists and describes a
number of safeguards and interlocks that must always be considered in connection with fire dilution ventilation. See also
Reference 10.5.
10.7
FIRE DILUTION VENTILATION FOR MIXTURES
It is common practice to regard the entire mixture as consisting of the components requiring the highest amount of dilution per unit liquid volume and to calculate the required air
quantity on that basis. This component would be the one with
the highest value for SG/[(MW)(LEL)].
10.8
VENTILATION FOR HEAT CONTROL
Ventilation for heat control in a hot industrial environment
is a specific application of general industrial ventilation. The
primary function of the ventilation system is to prevent the
acute discomfort, heat-induced illness, and possible injury of
those working in or generally occupying a designated hot
industrial environment. Heat-induced occupational illnesses,
injuries, or reduced productivity may occur in situations
where the total heat load may exceed the defenses of the body
and result in a heat stress situation. A heat control ventilation
system or other engineering control method must follow a
physiological evaluation in terms of potential heat stress for
the occupant in the hot industrial environment.
Due to the complexity of conducting a physiological evaluation, the criteria presented here are limited to general considerations. It is recommended, that the NIOSH Publication No.
86-113, Criteria for a Recommended Standard, Occupational
Exposure to Hot Environments,(10.6) be reviewed thoroughly in
the process of developing the heat control ventilation system.
The development of a ventilation system for a hot industrial
environment usually includes the control of the ventilation airflow rate, velocity, temperature, humidity, and airflow path
through the space in question. This may require inclusion of
certain phases of mechanical air-conditioning engineering
design which is outside the scope of this Manual. The necessary engineering design criteria that may be required are available in appropriate publications of the American Society of
Heating, Refrigerating and Air-Conditioning Engineers
(ASHRAE) handbook series.
10.9
HEAT BALANCE AND EXCHANGE
An essential requirement for continued normal body function is that the deep body core temperature be maintained
within the acceptable range of about 98.6 F [37 C] ± 1.8 F
[1 C]. To achieve an acceptable body temperature equilibrium
there must be a constant exchange of heat between the body
and the environment. The amount of heat exchanged is a function of 1) the total heat produced by the body (metabolic heat),
which may range from about 1 kilocalorie (kcal) per kilogram
(kg) of body weight per hour (1.16 watts) at rest to 5 kcal/kg
body weight/hour (7 watts) for moderately hard industrial
work; and 2) the heat gained, if any, from the environment.
The rate of heat exchange with the environment is a function
of air temperature and humidity, skin temperature, air velocity,
evaporation of sweat, radiant temperature, and type, amount,
and characteristics of the clothing worn, among other factors.
Respiratory heat loss is of little consequence in human defenses against heat stress.
The basic heat balance equation is:
DS = (M – W) ± C ± R – E
[10.15]
where: DS = change in body heat content
(M–W) = total metabolism – external work performed
C = convective heat exchange
R = radiant heat exchange
E = evaporative heat loss
General Industrial Ventilation
To solve the equation, measurement of metabolic heat
production, air temperature, air water vapor pressure, wind
velocity, and mean radiant temperature are required.
The major modes of heat exchange between man and the environment are conduction, convection, radiation, and evaporation.
10.9.1 Conduction. Other than for brief periods of body
contact with hot tools, equipment, floors, etc., which may
cause burns, conduction plays a minor role in industrial heat
stress. Because of the typically small areas of contact between
either body surfaces or its clothing and hot or cold objects, heat
exchange by thermal conduction is usually not evaluated in a
heat balance equation for humans. The effect of heat exchange
by thermal conduction in human thermal regulation is important only when large areas of the body are in contact with surfaces that are at temperatures different from average skin temperature (nominally 95 F [35 C]). It is important also when
even small body areas are in contact with objects that provide
steep thermal gradients for heat transfer.
In SI units, heat exchange is in watts per square meter of body
surface (W/m2). The heat exchange equations are available in
metric and English units for both the semi-nude individual and a
worker wearing conventional long-sleeved work shirt and
trousers. The values are in kcal/h or British thermal units per
hour (BTU/h) for the standard worker defined as one who
weighs 154 lbs [70 kg] and has a body surface area of 19.4 ft2
[1.8 m].
10.9.2 Convection. The rate of convective heat exchange
between the skin of a person and the ambient air immediately
surrounding the skin is a function of the difference in temperature between the ambient air (Ta), the mean weighted skin
temperature (Tsk) and the rate of air movement over the skin
(Va). This relationship is stated algebraically for the standard
worker wearing the customary one layer work clothing
ensemble as:
C = 0.65 Va0.6 (Ta – Tsk)
[10.16] IP
[C = 7 Va0.6 (Ta – Tsk)]
[10.16] SI
where: C = convective heat exchange, BTU/h [kcal]
Va = air velocity, fpm [m/s]
Ta = air temperature, F [C]
Tsk = mean weighted skin temperature,
usually assumed to be 95 F [35 C]
When Ta > 95 F [> 35 C], there will be a gain in body heat
from the ambient air by convection. When Ta < 95 F [< 35 C],
heat will be lost from the body to the ambient air by convection.
10.9.3 Radiation. Thermal radiant heat exchange between
the exposed surfaces of a person’s skin and clothing varies as
a function of the difference between the fourth power of the
absolute temperature of the exposed surfaces and that of the
surface of the radiant source or sink, the exposed areas and
10-17
their emissivities. Heat is gained by thermal radiation if the
facing surface is warmer than the average temperature of the
exposed skin and clothing and heat is lost by thermal radiation
if the facing surface is cooler than the average temperature of
the exposed skin and clothing. A practical approximation for
infrared radiant heat exchange for a person wearing conventional clothing is:
R = 15.0 (Tw – Tsk)
[10.17] IP
[R = 6.8 (Tw – Tsk)]
[10.17] SI
where: R = radiant heat exchange, BTU/h [kcal]
Tw = mean radiant temperature, F [C]
Tsk = mean weighted skin temperature, F [C]
10.9.4 Evaporation. The evaporation of water (sweat) or
other liquids from the skin or clothing surfaces results in a
heat loss from the body. Evaporative heat loss for humans is
a function of airflow over the skin and clothing surfaces, the
water vapor partial pressure gradient between the skin surface and the surrounding air, the area from which water or
other liquids are evaporating and mass transfer coefficients at
their surfaces.
0.6
E = 2.4 Va
(psk – pa)
[10.18] IP
0.6
[10.18] SI
[E = 110.4 Va
(psk – pa)]
where: E = evaporative heat loss, BTU/h [kcal]
Va = air velocity, fpm [m/s]
pa = water vapor pressure of ambient air, mmHg
psk = water vapor pressure on the skin, assumed to
be 42 mmHg at a 95 F [35 C] skin temperature
10.10 ACCLIMATIZATION OF THE BODY
Even people in generally good health can adjust physiologically to thermal stress only over a narrow range of environmental conditions. Diminished health status, medications,
limited prior thermal exposure, among other factors, increase
danger to thermal stresses.
People in general good health normally develop heat
acclimatization in a week or so after intermittently working or
exercising in a hot environment. Its effect is to improve the
comfort and safety of the heat exposure. Heat acclimatization
rapidly diminishes even after a day or so of discontinued activity in the heat — most is lost after about a week.
10.11 ACUTE HEAT DISORDERS
A variety of heat disorders can be distinguished clinically
when individuals are exposed to excessive heat. A brief
description of these disorders follows.
10.11.1 Heatstroke (also called Sunstroke). Heat stroke is
a life-threatening condition that, without exception, demands
10-18
Industrial Ventilation
immediate emergency medical care and hospitalization. Heat
stroke develops when body heat gains from exercise, work,
and/or a hot environment overwhelm normal thermoregulatory defenses.
black. Such a measure reports globe temperature (GT) (Figure
10-19). A person’s metabolic heat production is usually evaluated from an estimated level of average physical activity (Table
10-2).
10.11.2 Heat Exhaustion (also called Exercise-induced
Heat Exhaustion, Heat Syncope). Heat exhaustion most
Although there are a number of different indices for evaluating heat stress, none is reliable as a sole indicator of heat strain
for a specific person. Dry-bulb temperature is the least valuable
measure of heat stress because it provides no information about
ambient relative humidity, or heat exchange by convection or
radiation, and gives no estimate of the metabolic heat production. Wet-bulb globe temperature (WBGT) is often used as an
index of heat stress. When there is a source of radiant heat
transfer (solar radiation, hot surfaces of machinery):
commonly occurs in people who are not heat acclimatized
and who are in poor physical condition, obese, inappropriately dressed for a heat stress and exercising, or working energetically in the heat at unaccustomed and/or demanding tasks.
Although heat exhaustion is debilitating and uncomfortable,
it is not often a long-term health threat. There are considerable dangers, of course, for anyone operating machinery
when consciousness is impaired because of heat exhaustion
or for any other reason.
10.11.3 Heat Cramps (called Muscle Cramps) and Heat
Rash (called Prickly Heat, Miliaria Rubia). Spontaneous,
involuntary, painful, and prolonged muscle contractions commonly occur in otherwise healthy people when both body
water and electrolyte levels have not been restored after
extended periods of heavy sweating during exercise and/or
heat stress. Full recovery can be expected in about 24 hours
with the use of electrolyte replacement fluids and rest.
10.12 ASSESSMENT OF HEAT STRESS AND HEAT
STRAIN
WBGT = 0.7 Tnwb + 0.2 Tg + 0.1 Ta
[10.19]
where: Tnwb = natural wet-bulb temperature, F [C]
Tg = globe temperature, F [C]
Ta = ambient temperature, F [C]
When radiant heat transfer is negligible, Equation 10.19 is
replaced by:
WBGT = 0.7 Tnwb + 0.3 Tg
[10.20]
WBGT evaluates more factors contributing to heat stress
than does the measure of DB alone. It does not, however,
effectively evaluate the importance of mass and energy transfer from human skin by convection, which is essential for the
Heat stress is defined by environmental measurements of
air temperature, humidity, airflow rate, the level of radiant heat
exchange, and evaluation of a person’s metabolic heat production rate from exercise and/or work. Heat stress is the load on
thermoregulation. Heat strain is defined as the cost to each
person facing heat stress. Although all people working at the
same intensity in the same environment face the same level of
heat stress, each is under a unique level of heat strain. Because
disabilities, danger, and death arise directly from heat strain,
no measure of heat stress is a reliable indicator of a particular
person’s heat strain, or the safety of the exposure.
10.12.1 Evaluation of Heat Stress. Dry-bulb air temperature (DB) is measured by calibrated thermometers, thermistors,
thermocouples, and similar temperature-sensing devices,
which themselves do not produce heat and which are protected
from the effects of thermal conduction, evaporation, condensation, and radiant heat sources and sinks. Relative humidity is
evaluated psychrometrically as a function of the steady state
difference between dry-bulb temperature and that indicated by
the temperature of a sensor covered with a freely evaporating,
water-saturated cotton wick. Such a measure reports natural
wet-bulb temperature (NWB) when the wetted sensor is affected only by prevailing air movement, and wet bulb (WB) when
it is exposed to forced convection. Free air movement is measured with an unobstructed anemometer. Infrared radiant heat
transfer is typically measured by a temperature sensor at the
center of a 6-inch, hollow, copper sphere painted flat (matte)
FIGURE 10-19. Determination of wet-bulb globe temperature
General Industrial Ventilation
10-19
TABLE 10-2. Estimating Energy Consumed by Task/Work Performed
A. Body position and movement
Sitting
Standing
Walking
Walking uphill
B. Type of work
Hand work – light
Hand work – heavy
Work one arm – light
Work one arm – heavy
Work both arms – light
Work both arms – heavy
Work whole body – light
Work whole body – moderate
Work whole body – heavy
Work whole body – very heavy
BTU/hr [kcal/min]*
0.07 [0.3]
0.14 [0.6]
0.05–0.71 [2.0–3.0]
Add 0.06/ft rise [0.8 /meter rise]
BTU/hr [Average kcal/min]
Range
0.10 [0.4]
0.22 [0.9]
0.24 [1.0]
0.40 [1.7]
0.36 [1.5]
0.56 [2.5]
0.83 [3.5]
1.19 [5.0]
1.67 [7.0]
2.14 [9.0]
0.05–0.28 [0.2–1.2]
0.17–0.60 [0.7–2.5]
0.24–0.83 [1.0–3.5]
0.50–3.58 [2.5–15.0]
C. Basal metabolism
0.24 [1.0]
D. Sample calculation**
Assembling work with heavy hand tools
1. Standing
2. Two-arm work
3. Basal metabolism
TOTAL
0.14 [0.6]
0.83 [3.5]
0.24 [1.0]
1.24 BTU/hr [5.1 kcal/min]
*For standard worker of 154 lbs [70 kg] weight and body surface of 19.4 ft2 [1.8 m2].
**Example of measuring metabolic heat production of a worker when performing initial screening.
removal of heat from the skin surface and the formation of
water vapor from secreted sweat. Nor does WBGT evaluate
the importance of metabolic heat production in heat stress.
Under many environmental conditions, heat produced by
metabolism is the predominant, sometimes lethal, stressor.
quent training programs, and other information about
heat stress and strain.
6) Able to recognize the signs and symptoms of heat
strain in themselves and others exposed to heat stress
and know the appropriate steps for their remediation
(Figures 10-20 and 10-21).
10.13 WORKER PROTECTION
There is improved safety, comfort, and productivity when
those working in the heat are:
1) In generally good physical condition, not obese, heat
acclimatized, and experienced in the heat stressing job.
They also need to know how to select clothing and
maintain whole body hydration and electrolyte levels
to provide the greatest comfort and safety.
2) In areas that are well-ventilated and shielded from
infrared radiant heat sources.
3) Knowledgeable about the effects of their medications
affecting cardiovascular and peripheral vascular function, blood pressure control, body temperature maintenance, sweat gland activity, metabolic effects, and
levels of attention or consciousness.
4) Appropriately supervised when there is a history of abuse
or recovery from abuse of alcohol or other intoxicants.
5) Provided accurate verbal and written instructions, fre-
10.14 VENTILATION CONTROL
The control method presented here is limited to a general
engineering approach. Due to the complexity of evaluating a
potential heat stress producing situation, the accepted industrial hygiene method of recognition, evaluation, and control
should be utilized. In addition to the usual time limited exposures, it may be necessary to specify additional protection that
may include insulation, baffles, shields, partitions, personal
protective equipment, administrative control, and other measures to prevent possible heat stress. Ventilation control measures may require a source of cooler replacement air. Specific
guidelines, texts, and other publications or sources should be
reviewed for the necessary design information to develop the
ventilation system.
10.15 VENTILATION SYSTEMS
Exhaust ventilation can be used to remove excessive heat
and/or humidity if a replacement source of cooler and less
10-20
Industrial Ventilation
FIGURE 10-21. Recommended heat stress exposure limits
(acclimatized workers)
FIGURE 10-20. Recommended heat stress alert limits
(unacclimatized workers)
3
r = density of the air, lbm/ft [kg/m3]
cp = specific heat of the air, BTU/lbm-F
humid air is available. If it is possible to enclose the heat
source, such as in the case of ovens or certain furnaces, a gravity or forced air stack may be all that is necessary to prevent
excessive heat from entering the workroom. If a partial enclosure or local hood is indicated, control velocities, as shown in
Chapters 6 and 13, can be estimated from the volume of air to
be exhausted. It is important to remember that air entering the
enclosure may be at close to ambient conditions and might exit
at an elevated temperature. This resulting density change must
be considered when sizing the exhaust duct and fan.
Many operations do not lend themselves to local exhaust.
General ventilation may be the only alternative. The first step
in determining the required volumetric flow is to determine the
sensible and latent heat load. Next, determine the volumetric
flow to dissipate the sensible heat and the volumetric flow to
dissipate the latent heat. The required general ventilation is the
larger of the two volumetric flows.
The sensible heat rise can be determined by the following:
Hs = Qs H r H cp H DT H (60 min/hr)
[10.21] IP
[Hs = Qs H r H cp H DT H (60 s/min)]
[10.21] SI
where: Hs = sensible heat gain, BTU/hr [W]
Qs = volumetric flow for sensible heat,
acfm [am3/s]
[kJ/kg-C] or [kW-s/kg-C]
DT = Change in temperature, F [C]
For air, cp = 0.24 BTU/lbm-F [1.0 kW-s/kg-C] and r = 0.075
lbm/ft3 [1.204 kg/m3]. Consequently, the equation becomes:
Hs = 1.08 H Qs H DT
[Hs = 1.204 H Qs H DT]
[10.22] IP
[10.22] SI
or
Qs = Hs /(1.08 H DT)
[Qs = Hs /(1.204 H DT)]
In order to use this equation, it is necessary to first estimate
the heat load. This will include solar radiation, people, lights,
and motors as well as other particular sources of heat. Of these,
solar radiation, lights, and motors are all sensible sources. The
people heat load is part sensible and part latent. In the case of hot
processes that give off both sensible and latent heat, it will be
necessary to estimate the amounts or percentages of each. In
using the above equation for sensible heat, one must decide the
amount of temperature rise that will be permitted. Thus, in a
locality where 90 F [32 C] outdoor dry-bulb may be expected,
if it is desired that the inside temperature not exceed 100 F [38
C], or a 10 F [6 C] rise, a certain airflow rate will be necessary.
If an inside temperature of 95 F [35 C] is required, the necessary
airflow rate will be doubled.
General Industrial Ventilation
10-21
For latent heat load, the procedure is similar although more
difficult. If the total amount of water vapor is known, the heat
load can be estimated from the latent heat of vaporization, 970
BTU/lb (IP units). In a manner similar to the sensible heat calculations, the latent heat gain can be approximated by:
HL = QL H r H cL H Dω H (60 min/hr) H (1 lbm/7,000 grains) [10.23] IP
where: HL = latent heat gain, BTU/hr
QL = volumetric flow for latent heat, acfm
r = density of the air, lbm/ft3
cL = latent heat of vaporization, BTU/lbm
Dω = change in absolute humidity of the air,
grains-water/lbm-dry air
FIGURE 10-22. Good natural ventilation and circulation
For air, cL is approximately 970 BTU/lbm and r = 0.075
lbm/ft3. Consequently, the equation becomes:
HL = 0.62 H QL H Dω
[10.24] IP
or
QL = HL/(0.62 H Dω)
If the rate of moisture released (ṁ in pounds-mass per hour)
is known, then:
ṁ = QL H r H Dω H (1 lbm/7,000 gr) H (60 min/hr)
= QL H r H Dω/(0.00857)
HL = 45 × 103 × QL × Dw
or
or
QL = ṁ /(r H Dω H 0.00857)
ρ = density of the air, kg/m3
cL = latent heat of vaporization, W-s/kg
Δω = change in absolute humidity of the air,
kg water/kg dry air
For air, cL is approximately 2.256 × 106 W-s/kg and ρ = 1.2
kg/m3. Consequently, the equation becomes:
[10.25] IP
[10.26] SI
“Grains-water per pound-air difference” (ω) is taken from
the psychrometric chart or tables, and represents the difference
in moisture content of the outdoor air and the conditions
acceptable to the engineer designing the exhaust system. The
air quantities calculated from the above two equations should
not be added to arrive at the required quantity. Rather, the
higher quantity should be used since both sensible and latent
heat are absorbed simultaneously. Furthermore, in the majority
of cases, the sensible heat load far exceeds the latent heat load
so the design can be calculated only on the basis of sensible
heat.
The ventilation should be designed to flow through the hot
environment in a manner that will control the excess heat by
removing it from that environment. Figures 10-22 and 10-23
illustrate this principle.
In SI units, the procedure is similar. If the total amount of
water vapor is known, the heat load can be estimated from the
latent heat of vaporization, 2.256 × 106 W-s/kg. In a manner
similar to the sensible heat calculations, the latent heat gain
can be approximated by:
If the rate of moisture released (ṁ in kilograms per second)
is known, then
or
[10.27] SI
HL = QL × ρ × cL × Dw ) (60 s/min)
where:
HL = latent heat gain, Watts
QL = volumetric flow for latent heat, am3/s
FIGURE 10-23. Mechanically supplied ventilation
10-22
Industrial Ventilation
10.16 VELOCITY COOLING
If the air dry-bulb or wet-bulb temperatures are lower than
95–100 F [35–38 C], the worker may be cooled by convection
or evaporation. When the dry-bulb temperature is higher than
95–100 F [35–38 C], increased air velocity may add heat to
the worker by convection. If the wet-bulb temperature is high
also, evaporative heat loss may not increase proportionately,
and the net result will be an increase in the worker’s heat burden. Many designers consider that supply air dry-bulb temperature should not exceed 80 F for practical heat relief.
Current practice indicates that air velocities in Table 10-3
can be used successfully for direct cooling of workers. For best
results provide directional control of the air supply (Figure 1024) to accommodate daily and seasonal variations in heat
exposure and supply air temperature.
FIGURE 10-24. Spot cooling with volume and directional
control
10.17 RADIANT HEAT CONTROL
Since radiant heat is a form of heat energy that needs no
medium for its transfer, radiant heat cannot be controlled by
any of the above means. Painting or coating the surface of hot
bodies with materials having low radiation emission characteristics is one method of reducing radiation.
For materials such as molten masses of metal or glass that
cannot be controlled directly, radiation shields are effective.
These shields can consist of metal plates, screens, or other material interposed between the source of radiant heat and the workers. Shielding reduces the radiant heat load by reflecting the
major portion of the incident radiant heat away from the operator
and by re-emitting to the operator only a portion of that radiant
heat that has been absorbed. Table 10-4 indicates the percentage
of both reflection and emission of radiant heat associated with
some common shielding materials. Additional ventilation will
control the sensible heat load but will have only a minimal
effect, if any, upon the radiant heat load (Figure 10-25).
10.18 PROTECTIVE SUITS FOR SHORT EXPOSURES
For brief exposures to very high temperatures, insulated aluminized suits and other protective clothing may be worn.
These suits reduce the rate of heat gain by the body but provide
no means of removing body heat; therefore, only short exposures may be tolerated.
10.19 RESPIRATORY HEAT EXCHANGERS
For brief exposure to air of good quality but high temperature, a heat exchanger on a half-mask respirator face piece is
available. This device will bring air into the respiratory passages at a tolerable temperature but will not remove contaminants nor furnish oxygen in poor atmospheres.
TABLE 10-4. Relative Efficiencies of Common Shielding Materials
TABLE 10-3. Acceptable Comfort Air Motion at the Worker
Air Velocity, fpm* [m/s*]
Continuous Exposure
Air conditioned space
Fixed work station, general
ventilation or spot
cooling: Sitting
Standing
50–75 [0.25–0.38]
75–125 [0.38–0.63]
100–200 [0.50–1.00]
Intermittent Exposure, Spot Cooling or Relief Stations
Light heat loads and activity
Moderate heat loads and activity
High heat loads and activity
1,000–2,000 [5–10]
2,000–3,000 [10–15]
3,000–4,000 [15–20]
*Note: Velocities greater than 1,000 fpm [5 m/s] may seriously disrupt the
performance of nearby local exhaust systems. Care must be taken to direct air
motion to prevent such interference.
Surface of Shielding
Aluminum, bright
Zinc, bright
Aluminum, oxidized
Zinc, oxidized
Aluminum paint, new, clean
Aluminum paint, dull, dirty
Iron, sheet, smooth
Iron, sheet, oxidized
Brick
Lacquer, black
Lacquer, white
Asbestos board
Lacquer, flat black
Reflection of
Radiant Heat
Incident Upon
Surface
Emission of
Radiant Heat
from Surface
95
90
84
73
65
40
45
35
20
10
10
6
3
5
10
16
27
35
60
55
65
80
90
90
94
97
General Industrial Ventilation
10-23
ation is such that remote control is possible, an air conditioned
booth or cab can be utilized to keep the operator reasonably
comfortable.
10.22 INSULATION
If the source of heat is a surface giving rise to convection,
insulation at the surface will reduce this form of heat transfer.
Insulation by itself, however, will not usually be sufficient if
the temperature is very high or if the heat content is high.
REFERENCES
10.1
American Industrial Hygiene Association: The
Occupational Environment: Its Evaluation, Control &
Management, Second Edition (2003).
10.2
Air Force: AFOSH Standard 161.2 (1977).
10.3
American Conference of Governmental Industrial
Hygienists (ACGIH®): 2019 TLVs® and BEIs®
Book, Appendix E (2019).
10.4
National Fire Protection Association (NFPA) 86,
Standard for Ovens and Furnaces (2019).
10.5
Feiner, B.; Kingsley, L.: Ventilation of Industrial
Ovens. Air Conditioning, Heating and Ventilating,
pp. 82–89 (December 1956).
10.6
U.S. Department of Health and Human Services, PHS,
CDC, NIOSH: Occupational Exposure to Hot
Environments, Revised Criteria (1986).
FIGURE 10-25. Heat shielding
10.20 REFRIGERATED SUITS
Where individuals must move about, cold air may be blown
into a suit or hood worn as a portable enclosure. The usual
refrigeration methods may be used with insulated tubing to the
suit. It may be difficult, however, to deliver air at a sufficiently
low temperature. If compressed air is available, cold air may
be delivered from a vortex tube worn on the suit. Suits of this
type are commercially available.
10.21 ENCLOSURES
In certain hot industries, such as in steel mills, it is impractical to attempt to control the heat from the process. If the oper-
Supply Air Systems
11-1
Chapter 11
SUPPLY AIR SYSTEMS
NOTE: Equations with notation followed by (IP) are designated for inch-pound system only; equations followed by (SI) are designated for
metric use only. If equation bears neither, then it applies to both systems.
11.1 INTRODUCTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . .11-3
11.2 PURPOSE OF SUPPLY AIR SYSTEMS . . . . . . . . . . .11-3
11.2.1 Exhaust Air Replacement . . . . . . . . . . . . . . . . .11-3
11.2.2 Plant Ventilation . . . . . . . . . . . . . . . . . . . . . . . .11-4
11.2.3 Building Pressure . . . . . . . . . . . . . . . . . . . . . . .11-5
11.2.4 Building or Process Temperature Control,
Heating, and Cooling . . . . . . . . . . . . . . . . . . . .11-5
11.2.5 Product Protection and Space Air
Cleanliness . . . . . . . . . . . . . . . . . . . . . . . . . . . .11-7
11.3 SUPPLY AIR SYSTEM DESIGN FOR
INDUSTRIAL SPACES . . . . . . . . . . . . . . . . . . . . . . . . 11-7
11.3.1 General Manufacturing Areas . . . . . . . . . . . . .11-7
11.3.2 Shipping and Receiving Areas . . . . . . . . . . . . .11-9
11.3.3 Spaces with High Exhaust Volumes . . . . . . . .11-9
11.4 SUPPLY AIR EQUIPMENT . . . . . . . . . . . . . . . . . . . 11-11
11.4.1 Fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .11-11
11.4.2 Heating Systems . . . . . . . . . . . . . . . . . . . . . . .11-12
11.4.3 Steam Coil Heating . . . . . . . . . . . . . . . . . . . .11-12
11.4.4 Hot Water Coil Heating . . . . . . . . . . . . . . . . .11-14
11.4.5 Indirect Gas/Oil-fired Heating . . . . . . . . . . . .11-16
11.4.6 Direct Gas-fired Heaters . . . . . . . . . . . . . . . .11-16
11.4.7 Air Cooling Equipment . . . . . . . . . . . . . . . . .11-18
11.4.8 Mechanical Cooling . . . . . . . . . . . . . . . . . . . .11-19
11.4.9 Evaporative Cooling . . . . . . . . . . . . . . . . . . . .11-19
11.4.10 Air Filtration . . . . . . . . . . . . . . . . . . . . . . . . . .11-19
11.4.11 System Temperature Control . . . . . . . . . . . . .11-19
11.4.12 Unit Location . . . . . . . . . . . . . . . . . . . . . . . . .11-20
11.4.13 Size and Cost Considerations . . . . . . . . . . . . .11-20
11.5 SUPPLY AIR DISTRIBUTION . . . . . . . . . . . . . . . . . 11-20
11.5.1 Unidirectional or Plug Airflow . . . . . . . . . . .11-21
11.5.2 Mixing Ventilation Systems . . . . . . . . . . . . . .11-21
11.5.3 Air Displacement Ventilation Systems . . . . .11-22
11.5.4 Duct Materials . . . . . . . . . . . . . . . . . . . . . . . .11-23
11.5.5 Sheet Metal . . . . . . . . . . . . . . . . . . . . . . . . . . .11-23
11.5.6 Plastic . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .11-23
11.5.7 Fiberglass . . . . . . . . . . . . . . . . . . . . . . . . . . . .11-23
11.5.8 Textile . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .11-23
11.5.9 Supply Air System Design Considerations . .11-24
11.6 AIRFLOW RATE . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-24
11.6.1 Air Changes . . . . . . . . . . . . . . . . . . . . . . . . . .11-24
11.7 HEATING, COOLING AND OTHER
OPERATING COSTS . . . . . . . . . . . . . . . . . . . . . . . . . 11-25
11.7.1 Estimating Heating Energy Use . . . . . . . . . . .11-25
11.7.2 Air Supply vs. Plant Heating Costs . . . . . . . .11-26
11.7.3 Energy Considerations . . . . . . . . . . . . . . . . . .11-26
11.7.4 System Maintenance . . . . . . . . . . . . . . . . . . .11-26
11.7.5 Untempered Air Supply . . . . . . . . . . . . . . . . .11-26
11.7.6 Energy Recovery . . . . . . . . . . . . . . . . . . . . . .11-26
11.8 INDUSTRIAL EXHAUST RECIRCULATION . . . . 11-26
11.8.1 Evaluation of Employee Exposure Levels . . .11-27
11.8.2 Design Considerations for Air
Recirculation . . . . . . . . . . . . . . . . . . . . . . . . .11-29
11.8.3 Recirculation Air Monitor Selection . . . . . . .11-29
11.9 SYSTEM CONTROL . . . . . . . . . . . . . . . . . . . . . . . . 11-30
11.9.1 Building Air Balance . . . . . . . . . . . . . . . . . . .11-30
11.9.2 Temperature . . . . . . . . . . . . . . . . . . . . . . . . . .11-30
11.9.3 Indoor Air Quality . . . . . . . . . . . . . . . . . . . . .11-31
11.10 SYSTEM NOISE . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-31
REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .11-32
____________________________________________________________
Figure 11-1
Cold Zones vs. Overheated Zones (Poor
Ventilation Design) . . . . . . . . . . . . . . . . . . . . . 11-4
Figure 11-2 (IP) Relationship Between Air Pressure and
Amount of Force Needed to Open or
Close an Average-sized Door . . . . . . . . . . . . . 11-5
Figure 11-2 (SI) Relationship Between Air Pressure and
Amount of Force Needed to Open or
Close an Average-sized Door . . . . . . . . . . . . . 11-5
Figure 11-3
How Fan Performance Decreases with
Negative Pressure . . . . . . . . . . . . . . . . . . . . . . 11-6
Figure 11-4
Figure 11-5
Figure 11-6
Figure 11-7
Figure 11-8
Figure 11-9
Figure 11-10
Figure 11-11
Figure 11-12
Figure 11-13
Types of Supply Air System Designs . . . . . . .11-8
Types of Door Heater Designs . . . . . . . . . . .11-10
Direct-fired Unit . . . . . . . . . . . . . . . . . . . . . . 11-11
Single Steam Coil Unit . . . . . . . . . . . . . . . . .11-14
Steam Coil . . . . . . . . . . . . . . . . . . . . . . . . . . .11-15
Multiple Coil Steam Unit . . . . . . . . . . . . . . . 11-16
By-pass Steam System . . . . . . . . . . . . . . . . .11-16
Integral Face and By-pass Coil . . . . . . . . . . .11-17
Indirect-fired Unit . . . . . . . . . . . . . . . . . . . . .11-17
Direct-fired By-pass Unit . . . . . . . . . . . . . . .11-17
11-2
Industrial Ventilation
Figure 11-14
Figure 11-15
Figure 11-16
Figure 11-17
Air Heating and Cooling Requirements . . . .11-20
Air Jet Temperature and Velocity
Profile (IP Units) . . . . . . . . . . . . . . . . . . . . . .11-21
Airflow in Displacement Ventilation
System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-23
Register Airflow Patterns . . . . . . . . . . . . . . . 11-25
Figure 11-18
Figure 11-19
Figure 11-20
Recirculation Decision Logic . . . . . . . . . . . . 11-27
Schematic Diagram of Recirculation
Monitoring System . . . . . . . . . . . . . . . . . . . . 11-30
Schematic of Recirculation from Air
Cleaning Devices (Particulates) . . . . . . . . . . 11-31
____________________________________________________________
Table 11-1
Table 11-2
Table 11-3
Negative Pressures That May Cause Unsatisfactory Conditions within Buildings . . . . . . . .11-4
Negative Pressures and Corresponding
Velocities through Crack Openings . . . . . . . . .11-4
Summary of Advantages and Limitations
of Typical Industrial Heating Sources . . . . . .11-13
Table 11-4
Comparison of Heater Advantages and
Disadvantages . . . . . . . . . . . . . . . . . . . . . . . . .11-18
Table 11-5 (IP) Air Exchanges vs. Room Size . . . . . . . . . . . .11-24
Table 11-5 (SI) Air Exchanges vs. Room Size . . . . . . . . . . . .11-24
Supply Air Systems
11.1
INTRODUCTION
Industrial buildings operating in the early 1900s had simple
building mechanical systems. Ventilation was accomplished
by opening a wall/roof section and letting the outside air naturally flow through the building. Heating systems consisted of
radiators and unit heaters. As more automation was incorporated into the industrial process, buildings had to deal with
increasing amounts of energy being consumed inside. Some
process operations created potentially hazardous emissions in
the worker’s environment. This caused the need to install
exhaust air systems to control these airborne emissions. With
the use of powered exhaust systems, many buildings began to
operate with a negative pressure. Supply air equipment was
soon found to be critical to the success of industrial ventilation
systems. They provide the air that allows exhaust systems to
perform properly. In some situations, they also provide dilution of contaminants that escape into the general workspace.
Over the years, heating and ventilating units advanced to provide a more comfortable building temperature at a lower energy use when compared to a system that uses only unit heaters.
Manufacturing facilities evolved to the point where there is
now widespread use of automation/computers. Production of
parts requiring tight tolerances is often required. These facilities require temperature control to perform at effective levels.
Workers need to be cooled to relieve body heat caused by their
activity. This heat exchange is easily accomplished with cool
air. With warmer temperatures that occur in the summer or
near hot industrial operations, maintaining a suitable rate of
cooling becomes more difficult. Refer to Chapter 10 for more
information regarding worker cooling.
In some industrial plants, ventilation systems are key elements of a process. In a few it is critical to the success of that
process; this is the case in automotive painting. A number of
years ago, automobiles were painted in an open booth by people who sprayed paint onto the vehicle body. Air was exhausted to remove solvent vapors so hazardous conditions and
explosive concentrations would not exist. The replacement
make-up air entering the booth had a minimum degree of filtration and no significant temperature or humidity control.
Supply air was distributed to provide good air exchange
throughout the booth so the concentration of paint solvent
vapors would be low. Over time, the quality of the paint coating became more important and the performance of ventilation
systems began to improve. Currently supply air humidity and
temperature are controlled to improve paint curing time. The
air is well filtered to eliminate defects in the painted surface.
Painting operations are conducted in a clean-room type space
that is pressurized to maintain high levels of cleanliness.
In other plants, the distribution of supply air may not be as
critical for product quality but will always be important for the
proper operation of exhaust systems and plant comfort control. Poorly distributed supply air sometimes overwhelms a
well-designed exhaust hood and impedes the hood’s ability to
capture contaminants. Therefore, the designer should pay
11-3
equal attention to both the quantity and distribution of the supply air system.
11.2
PURPOSE OF SUPPLY AIR SYSTEMS
A proper supply air ventilation system can serve several
purposes in an industrial facility: 1) exhaust air replacement,
2) plant ventilation, 3) building pressurization, 4) building
heating, cooling, and humidification, and 5) space air cleanliness. The purposes of the supply air system are discussed in
the following paragraphs. The total amount of supply air
should be the amount that satisfies all the requirements of the
supply air system. For example, a small amount of air may be
required for replacing the exhaust air, but a much larger
amount may be required to deliver enough tempered air for
heating or cooling.
11.2.1 Exhaust Air Replacement. Air will enter a building
in an amount equal to the flow rate of exhaust air whether or
not provision is made for replacement supply air systems or
building infiltration. However, the actual exhaust flow rate will
be less than the design value if the plant is under negative pressure. With inadequate supply air and if the building perimeter
is tightly sealed, blocking effective infiltration of outdoor air, a
severe decrease of the exhaust flow rate will result. If, on the
other hand, the building has large sash areas, air infiltration
may be quite pronounced and the exhaust system performance
will decrease only slightly. This situation may cause other
problems as identified in Table 11-1 where the resulting inplant environmental condition is often undesirable since the
influx of cold outdoor air in the northern climates chills the
perimeter of the building. Exposed workers are subjected to
drafts, space temperatures are not uniform, and the building
heating system is usually overtaxed (Figure 11-1). Under such
negative pressure conditions, workers in the cold zones turn up
thermostats in an attempt to get heat. Because this will do nothing to stop leakage of cold air, they remain cold while the center of the plant is overheated. Although the air may eventually
be tempered to acceptable conditions by mixing with warmer
air as it moves to the building interior, this is an ineffective way
of transferring heat to the air and usually results in fuel waste.
For an estimated value of the airflow entering a building
through cracks occurring around doors or windows or other
small openings in a building exterior, refer to Table 11-2.
Figure 11-2 presents the force necessary to open a door against
a building’s negative pressure. The performance of an exhaust
fan operation can also suffer as shown in Figure 11-3.
For general plant ventilation, replacement airflow rate
should be slightly more than the total airflow rate removed
from the building by exhaust ventilation systems, process systems, and combustion processes. Determination of the actual
flow rate of air removed usually requires an inventory of
exhaust locations with airflow testing of these sources. When
conducting the exhaust inventory, it is necessary not only to
determine the quantity of air removed, but also to identify the
11-4
Industrial Ventilation
TABLE 11-1. Negative Pressures That May Cause Unsatisfactory Conditions Within Buildings
Negative Pressure, "wg [Pa]
Adverse Conditions
0.01 to 0.02 [3–5]
Worker Draft Complaints—High velocity drafts through doors and windows.
0.01 to 0.05 [3–13]
Natural Draft Stacks Ineffective—Ventilation through roof exhaust ventilators, flow through stacks with natural
draft greatly reduced.
0.02 to 0.05 [5–13]
Carbon Monoxide Hazard—Back drafting will take place in hot water heaters, unit heaters, furnaces, and
other combustion equipment not provided with induced draft fan.
0.03 to 0.10 [8–25]
General Mechanical Ventilation Reduced—Airflows reduced in propeller fans and low pressure supply and
exhaust systems.
0.05 to 0.10 [13–25]
Doors Difficult to Open—Serious injury may result from non-checked, slamming doors.
0.10 to 0.25 [25–63]
Local Exhaust Ventilation Impaired—Centrifugal fan exhaust airflow reduced.
need to upgrade any part of the ventilation system. At the
same time, reasonable projections should be made of the total
plant exhaust requirements for the next few years, particularly
if process changes or plant expansions are contemplated. In
such cases it can be practical to purchase a replacement air
unit slightly larger than immediately necessary with the
knowledge that the increased capacity will be required within
a short time. The additional cost of a larger unit is relatively
small and, in most cases, the fan drive can be adjusted to supply the desired quantity of air at the time of installation.
Having established the minimum air supply quantity necessary for replacement air purposes, many plants have found that
it is wise to provide additional supply airflow to overcome natural ventilation leakage and further minimize drafts at the
perimeter of the building. Conversely, some facilities deliberately design for a higher exhaust flow rate to prevent fugitive
emissions from migrating into “clean” areas of the building or
to the outdoors. In these situations, the control of the building
pressure is quite important.
11.2.2 Plant Ventilation. Outside air brought into an indus-
trial plant is utilized to replace air exhausted, and may help
FIGURE 11-1. Cold zones vs. overheated zones (poor ventilation design)
dilute airborne contaminants present in the workspace. As discussed in Chapters 3, 4, 6 and 9, exhaust air systems are used
to remove unwanted airborne contaminants, heat, odors, and
gases by placement as close to the source of generation as possible. The supply air system can aid in contaminant control by
diluting remaining contaminants in the general workspace
with outdoor air. Chapter 10 discusses the design approach for
sizing the supply air rate for this purpose. Outdoor air can also
be used to reduce the building’s temperature by blending the
TABLE 11-2. Negative Pressures and Corresponding Velocities
Through Crack Openings (Calculated with air at room temperature,
standard atmospheric pressure, Ce = 0.6.)
Negative Pressure, "wg [Pa]
Velocity, fpm [m/s]
0.004 [1]
150 [0.75]
0.008 [2]
215 [1.08]
0.010 [3]
240 [1.20]
0.016 [4]
300 [1.50]
0.020 [5]
340 [1.70]
0.025 [6]
380 [1.90]
0.030 [8]
415 [2.08]
0.040 [10]
480 [2.40]
0.050 [13]
540 [2.70]
0.060 [15]
590 [2.95]
0.080 [20]
680 [3.40]
0.100 [25]
760 [3.80]
0.150 [38]
930 [4.65]
0.200 [50]
1080 [5.40]
0.250 [63]
1200 [6.00]
0.300 [75]
1310 [6.55]
0.400 [100]
1520 [7.60]
0.500 [125]
1700 [8.50]
0.600 [150]
1860 [9.30]
Supply Air Systems
FIGURE 11-2 (IP). Relationship between air pressure and
amount of force needed to open or close an average-sized
door.
warmer plant air with cooler outside air. The air can be blown
across a person to achieve a greater cooling effect than still air.
Chapter 10 also discusses heat relief and measurements related to cooling for occupant comfort.
Ventilation air is also needed to deliver oxygen for breathing. This is a concern with confined spaces, but most industrial plants have somewhat porous building shells, and outside
air infiltration is normally more than adequate to provide fresh
air for breathing. Air can easily flow through cracks around
doors, operable windows, utility entrances, conveyor openings, and through roof mounted equipment components.
Infiltration of air in this manner is not a substitution for outdoor air provided by a supply air system and may not satisfy
the requirements of ASHRAE 62.1, Ventilation for Indoor Air
Quality.
FIGURE 11-2 (SI). Relationship between air pressure and
amount of force needed to open or close an average-sized
door.
11-5
11.2.3 Building Pressure. While negative pressure can
cause adverse conditions, there are situations where negative
pressures are desired. An example is a room or area where a
contaminant must be prevented from escaping into the surrounding area. It may also be desirable to maintain a room or
area under positive pressure to maintain a clean environment.
Either of these conditions can be achieved by setting and
maintaining the proper exhaust/supply flow differential.
Negative pressure can be achieved by setting the exhaust volumetric flow rate (Q) from the area to a level higher than the
supply rate. A good performance standard for industrial processes is to set a negative pressure differential of 0.04 ± 0.02
"wg [10 ± 5 Pa]. Conversely, positive pressure is achieved by
setting the supply airflow rate higher than the exhaust rate. The
proper flow differential will depend on the physical conditions
of the area, but a general guide is to set a 5% flow difference
but no less than 50 acfm [0.03 am3/s]. If the volume flows vary
during either a negatively or positively pressurized process, it
is easier to maintain the desired room pressure by adjusting the
supply air.
11.2.4 Building or Process Temperature Control, Heating,
and Cooling. In addition to contaminants, which are most
effectively controlled by hoods, industrial processes may create
an undesirable heat load in the workspace. Modern automated
machining, conveying, and transferring equipment requires
considerable horsepower. It is not uncommon for the process to
have an electrical use of 10 to 20 watts per square foot [110–220
watts/m2] of floor space. This equals a heat input of 34 to 68
BTU per hour in IP units. Precision manufacturing and assembling demand increasingly higher light levels in the plant with
correspondingly greater heat release. The resulting in-plant heat
release raises indoor temperatures, at times beyond the limits of
efficient and healthful working conditions and, in some cases,
beyond the tolerance limits for the product.
Environmental control of these factors can be accommodated through the careful planning and use of the supply air system. Industrial air conditioning may be required to maintain
process specifications and reduce hot working conditions.
For a large industrial plant whose size is several hundred
thousand square feet, the internal process heat may more than
equal the heat loss through the building’s walls and roof on the
coldest of days. Therefore, this plant needs to be cooled
throughout the year. The supply air must be heated to the
degree that cold drafts are avoided. Heated air should also be
utilized at door openings to reduce the cold drafts occurring
with an open door. With these large facilities, the issue is how
best to accomplish plant cooling.
The engineer in charge of providing suitable in-plant temperatures must understand and consider the building occupant
needs as well as those of the building. Humans must lose heat
to survive and lose it at a controlled rate to be comfortable.
Therefore, the design engineer who is trying to achieve human
comfort sometimes has a heating concern, but always has a
cooling problem.
11-6
Industrial Ventilation
Supply Air Systems
Cooling the workspace in the summer is often more difficult
than heating this space. In the heating season, the outdoor air
temperatures are cool and it is relatively easy to obtain a 60 F
to 70 F [16 C to 21 C] supply air temperature that provides a
suitable workspace environment with normal process heat
release to the space. In the summer when the outside temperature is in the 80s and 90s [approx. 27 C to 30 C], reasonable
space temperatures can be obtained by bringing in additional
outside air, increasing the air velocity over the person, or using
evaporative coolers/refrigeration equipment to cool the supply
air.
When applying a cooling system to industrial operations, a
common objective is to obtain a plant temperature of approximately 80 F [27 C]. The intent is not to try to provide a high
level of comfort or control humidity; it is only to control heat.
ASHRAE(11.1) gives basic criteria for industrial air conditioning
in HVAC applications. Sensible and latent heat released by people and processes can be controlled to desired limits by proper
use of air conditioning equipment. Radiant heat cannot be controlled by cooler air or increased ventilation, thus methods such
as shielding, described in Chapter 10, are required.
To obtain the most cost-effective cooling system, a comparison should be performed between the use of extra air and
cooling the air. The extra air approach often uses significantly
more than the required wintertime airflow of outside air to
dilute the process heat for adequate workspace temperature.
This results in the operation of an oversized fan and air distribution system during most of the year. Compare this to a winter-sized system with cooling capability that is used only when
needed.
For industrial plants larger than 400,000 square feet [36,000
m2] in size, the supply inlet air temperature in a ventilation
system is typically 5 to 10 degrees F [3 C to 6 C] warmer in
the summer than the actual outside temperature. The combination of the building process heat and solar radiation heat on the
roof results in the situation that the air taken into a rooftop supply unit intake is warmer than the surrounding ambient air. To
minimize this temperature increase, the unit’s outside air
intake opening should be a distance above the roof that is equal
to the sum of at least two feet plus the effective diameter of the
intake opening. The fan and motor also increase the air temperature by approximately three degrees. If the motor is located outside the air stream, the temperature rise can be reduced
by two degrees F [1 C].
11.2.5 Product Protection and Space Air Cleanliness. If a
space requires a higher level of cleanliness than adjacent spaces,
there should be an excess flow of clean air into the clean space,
resulting in space pressurization and an outward airflow from
the clean space to the less clean spaces. The clean air displaces
the air in the space and the amount of airborne contaminants is
reduced. To achieve a high degree of air cleanliness, special filters provide the final filtration. Refer to Chapter 8 for air cleaning characteristics of HEPA and other filter systems. The air
exchange rate of cleanrooms must increase to achieve higher
11-7
degrees of air cleanliness depending upon the process and work
practices involved. It is important to balance the need for product cleanliness and worker protection. In most situations, the
supply air enters the room from ceiling panels or diffusers and
is exhausted near the floor. Room velocities in the range of 50
to 100 feet per minute are typically used in the cleanest spaces.
Filters for typical commercial buildings have a minimum efficiency reporting value (MERV) around 7 or 8 based on the
ASHRAE Standard 52.2, Method of Testing General Ventilation
Air Cleaning Devices for Removal Efficiency by Particle
Size.(11.4) The MERV indicates a filter’s initial efficiency as a
function of particle size, and gives a numeric value that allows
a user or engineer to specify the appropriate rating. When a process requires a high level of cleanliness, such as food processing, painting, or assembly of parts where a fine dust is a detriment, a more efficient filtration system is required. Facilities
such as hospitals and other clean operations may need filters
with a MERV rating as high as 13 to 16. Refer to Chapter 8 for
more discussion regarding air cleaning equipment.
11.3
SUPPLY AIR SYSTEM DESIGN FOR
INDUSTRIAL SPACES
The design of the supply air system must satisfy several
requirements for success. The air must enter the space without
disturbing the performance of local exhaust systems or process
equipment operation and without causing undesired drafts or
excessive noise. High velocity airflows created by large volumes of supply air directed out of supply air registers can ruin
the effectiveness of a local exhaust system. Processes involving powders, extrusions of thin membranes or the handling of
objects easily dislodged by air movement are not tolerant of
high velocity air streams. Employees who are reasonably comfortable often dislike high air velocities that result in unwanted
drafts. The movement of high velocity air through the supply
air system can also result in objectionable noise.
There are several types of spaces that occur in an industrial
facility that require care in the design of supply air systems.
They are discussed in the following paragraphs.
11.3.1 General Manufacturing Areas. This space is an open
area with the process equipment and people spread throughout. The processes may or may not have local exhaust systems
associated with them. The purpose of the supply air system is
to provide exhaust replacement air, general ventilation, and
temperature control. Several approaches to the supply air system design are shown in Figure 11-4. The ventilation system
choices shown include the use of unit heaters that provide no
ventilation and are poor at controlling the space temperature.
With this choice, make up air enters the building by the infiltration of unconditioned outside air through doors, windows,
and other openings. The only means of adjusting the rate of
airflow is by opening or closing windows or some other building elements open to the outside. This type of system can produce cold drafts that cause employee complaints. It also has
high energy usage since the cold drafts make the heating sys-
11-8
Industrial Ventilation
Supply Air Systems
tem work harder. This type of system often provides an
uncomfortable plant interior since the hard working heating
system raises the temperature in this area to excessive temperatures.
Another type of system (high level ventilation) is one that
can bring in outside air, heat that air, and deliver it into the
building using minimal air distribution ducts. This system has
the ability to provide reasonably good thermal conditions during the heating season, but when the supply air is warmer than
the air at the floor level, very little of the ventilation air is able
to enter the worker zone. This system can be a poor replacement air system for local exhaust. The large mass of air
released into the space causes high air velocities that can
adversely affect the performance of exhaust hoods. Also, many
processes release heat as they operate, causing the surrounding
air to become warmer and forcing it to rise. As this air moves
upward, it carries contaminants expelled by the process. By
introducing the supply air in the truss space, the fresh air mixes
with the process contaminants often resulting in the contaminants being pushed back down to the workers.
A third system is similar to the second system except that
there is a ducted air distribution system. The results are similar
to the second system since the air is still discharged in the
upper level of the plant; and, like the second system, when the
supply air is warmer than the space air, it will stay above the
occupied zone unless it is forced down with high velocity outlets. Registers used as air outlets will entrain room air into the
supply air stream.
A fourth design (low level distribution) is the same as the
third, except the supply air outlets are dropped to the worker
level. The purpose of this design is to place the air discharge
low enough to provide a cooler air temperature and not disturb
the warmer air located in the truss space. Good design for maintaining cool summer temperatures is to have a system that has
the air discharged at approximately 8 to 10 feet [2.4 to 3.0 m] off
of the floor.
When an older plant is being renovated for improved ventilation, a system for the entire plant should be considered. The
system does not need to be installed at one time and can be
constructed in phases. Since the ventilation air mixes readily
with the plant air, the need to treat areas or spots in a manufacturing plant is normally not needed. This allows the installation of a repetitive system design without a significant reduction in performance. Using the repetitive system approach provides a lower cost system since many of the components can
be duplicated. Each duct system register box and registers
should be the same. The use of common devices simplifies the
maintenance of the system and provides a more flexible system to operate.
11.3.2 Shipping and Receiving Areas. Plants looking to
upgrade their ventilation system should first consider providing adequate door heaters or air curtains at their primary outside truck doors. For plants that have a negative pressure con-
11-9
dition, the doors will be a source of cold drafts and lost building heat. The design of a door heater can take one of three
approaches as shown in Figure 11-5. The first type of door
heater uses air that is discharged at the top of the door blowing
down over the opening. This type works reasonably well if the
door is not too high (12 feet [3.6 m] or less) and the plant’s
building pressure is neutral or positive. The second type has a
duct system that directs the heated air horizontally across the
door opening from each side. This approach works better for
taller doors since the throw of air is shorter and much of the
heated air is provided close to the floor. The delivered air can
readily mix with the incoming cold air that is dense and wants
to flow along the floor. The third option is significantly more
costly than the first two and should be reserved for very large
doors. This door heater delivers the air through an opening
that is in the floor running the width of the door. Since truck
traffic using the door will ride over the opening, it must be
covered with steel grating that can support the vehicle and
allow the heated air to be blown through it. This type of door
heater provides the best results since the warm air is blown
upward at the door opening where it mixes with the incoming
cold air warming the cold draft. With the air being warmed at
the lowest elevation, people in the occupied zone get full benefit of the heat provided.
The required airflow for the proper performance of these
door heaters is dependent on the amount of negative pressure
in the building, the wind force commonly present and the outside temperature. Typically, a value of 100 acfm per square
foot [0.5 am3/s/m2] of door opening is utilized for door heater
sizing in a building with a neutral or positive pressure. The
discharge air velocity at the outlet should be approximately
3,000 feet per minute [15.0 m/s] for those heaters that discharge air down or from the sides. The door heater type that
discharges air up requires a lower air velocity, generally less
than 1,000 feet per minute [50 m/s].
11.3.3 Spaces with High Exhaust Volumes. Some spaces
require large quantities of make-up air to satisfy the exhaust
airflow requirement. A major issue is how to introduce the air
into the room without adversely affecting the performance of
those exhaust systems. Significant air velocities across the face
of a hood can greatly affect its performance to capture the contaminants it was installed to control. In this type of space, supply air should be released at low velocities. One option is the
use of a perforated duct that has a number of openings through
which air is released into the room at low velocities. An alternate approach would be to use a plenum with perforated sides
or bottom to release the air with little velocity. Both of these
approaches work well in small spaces that have high levels of
exhaust. The plenum or perforated duct should be placed
behind or above the workers so that clean air movement is
over or behind them on the way to the exhaust in enclosures
such as paint booths, small laboratories, fiberglass lay up and
spray up rooms, etc.
In large spaces, the supply air can be released in any manner
11-10
Industrial Ventilation
Supply Air Systems
that does not cause excessive air movement near the exhaust
hoods. Care should be taken to assure that the discharge air is
not hitting vertical surfaces and creating unwanted high velocities in the occupied level.
11.4
SUPPLY AIR EQUIPMENT
A supply ventilation system consists of the supply air handling unit, the air distribution duct, and the supply air outlet.
The supply air unit has components to temper, clean, and
move the air. A microprocessor normally controls these
devices through sensors and actuators. Since there can be significant internal heat generation in many industrial plants,
space cooling is the objective for most of the year.
There are several grades of air handling units: heavy industrial, light industrial, and commercial. The heavy industrial
units are normally a custom or modular type and can provide
many years of continuous service. If well maintained, they can
easily operate for 20 years or more. The components are
stronger and there is significantly more space for access to
fans, filters, coils, and dampers. This facilitates the ability to
maintain and repair the equipment. A light industrial grade unit
typically provides less space for components maintenance
making it more difficult to change belts, motors, etc. Parts are
less suitable for rugged industrial use. They are often massproduced with some flexibility to make modifications. They
offer the same wide choice of heating and cooling media as the
heavy industrial unit. In contrast, the commercial unit has less
of a choice of heating and cooling equipment, is mass produced, has a minimum amount of space for maintenance, and
is structurally designed for non-industrial buildings.
Unit heaters and fan coils (Figure 11-6) are also utilized in an
industrial space. The unit heater is a low cost, heating-only unit.
It uses a propeller fan to push room air through a heating coil or
fuel-fired furnace. It is used for spot heating since each unit has
a limited capacity. Typically, unit heaters are hung from the
building structure and located to blow into a specific area. Fan
coil units are similar in function except they are normally placed
against a wall at floor level. They are most often found at building entrances, administrative areas, and similar spaces.
Most air handling units are manufactured in a factory and
shipped to the industrial site. Years ago, it was common to
have units that were erected in the field. The fan, coils, filters
FIGURE 11-6. Direct-fired unit
11-11
and other components were delivered to the site and installed
in a sheet metal enclosure that formed the air handling unit.
Penthouses would be erected on the roof to house these air
handling units. Today, field erected units are too expensive
when compared to factory built units and are used only for
special applications. Large factory units are designed to be
split into a number of sections sized for ease of handling and
shipping to the site. When installing these units at a large manufacturing plant, it is common to lift the air handling sections
into place by helicopters. Most helicopters used for this purpose have a maximum lifting capability of approximately
8,000 to 10,000 pounds [3600 to 4500 kg]. Another consideration of air handling unit size is the dimension restrictions
regarding over the road travel. The maximum trailer width is
typically 12 feet [3.6 m] and the normal height limit is 13.5
feet [4.05 m] off the road with the unit sitting on the trailer.
Factory manufactured air handling units are completely
assembled and tested before they leave the plant. After they
have passed the necessary tests and approvals, they are disassembled and made ready for shipment. These units are constructed so that sections can be unbolted from each other. The
piping has joints to allow quick disassembly. The electrical
wiring has junction boxes near the joints or wiring is pulled
back from one of the connection points. Units as large as
100,000 to 150,000 acfm [50 to 75 am3/s] capacity are constructed in this manner. When they are received at the construction site, they are lifted into place and all sections are reattached to form the air handling unit. The electrical wiring is
reinstalled along with the necessary piping connections.
Having tested the unit in the factory, equipment startup usually
goes smoothly.
Units that are commercial grade are normally shipped in
one section and this limits their size. Their size is also limited
by the sales demand. Since they are mass produced on an
assembly line, significant demand is required to warrant production of a particular size unit. As a result, the highest volume
units are in the size ranges below 40,000 acfm [20 am3/s].
11.4.1 Fans. The heart of the air handling unit is the fan. It is
the device that causes air to flow through the supply air system.
To size the fan properly, the quantity of airflow must be identified as well as the static pressure loss due to the elements in the
system that resist flow. The air quantity is determined by the purpose of the system. If the unit provides makeup air for an
exhaust system, the airflow quantity depends on the system use
as discussed in Section 11.2. If the system is to provide heating
and/or cooling for a building, then the airflow becomes the
quantity required to satisfy the heating/cooling load. Once the
total airflow of the building space is identified and the number
of units is chosen, the airflow required for each unit can be determined. The static pressure required for system flow is determined similar to the way exhaust system static pressure is calculated.
The fan selection for a supply fan is the same as that used to
select an exhaust fan. The types of fans used are different since
11-12
Industrial Ventilation
the static pressure is normally lower than that of an exhaust
system and the air is cleaner. Common fans associated with
ducted supply air systems are forward curved or backward
inclined blade centrifugal fans. These fans have the capability
to generate several inches of static pressure needed to move
the air through the air handling unit and duct distribution system. In some units a plenum fan is used. This is a centrifugal
fan wheel placed in a sheet metal enclosure. The rotating fan
wheel pressurizes the enclosure and openings are made in the
enclosure for airflow out. The use of a plenum fan can eliminate the need for an elbow near the air discharge, but some loss
in efficiency occurs.
cooling (when moisture is removed) takes place. The latent
cooling energy use for a system can be significant. The energy
is required to condense the water vapor in the supply air stream
when the desired cool temperature is below the dew point of
the air. Thus as the supply air is cooled, the moisture content
of the air reaches the temperature when the air has a 100% relative humidity. At this point, to achieve a lower supply air temperature, moisture in the air begins to condense on the cooling
coil. This latent cooling energy use can be determined by the
formula:
When selecting a fan, choose one that can be upgraded to
meet more demanding operating conditions. This will give the
system the flexibility to meet future needs. Fans are built to
achieve different levels of service (Class I to Class IV). The
Class IV fan is designed to be strong enough to handle the
stresses of the highest fan outlet velocity and pressure. When
selecting a fan for a range of service, the fan laws must be considered to understand the limitations for varied flow. It is customary to select a fan that will operate at no more than 80% of
its full rated speed. The motor selected should be able to provide the horsepower required to achieve that full speed. The
electrical service for the fan should be designed to handle the
horsepower [watts] required for the speed increase of 20%.
The motor horsepower goes up as the cube of the increase in
speed. Be sure to have the power required for a cold start of
the fan, even if it is to operate continuously. All fans need to
be shut down for maintenance. Refer to Chapter 7 in this
Manual and the ASHRAE Handbooks for more information
regarding fan selection.
where: 1,076 BTU/lbm is the approximate heat content of a
50% relative humid vapor at 75 F minus the heat content of 50
F water, the normal cooling coil condensing temperature. For
the total cooling energy use, the sensible cooling load needs to
be added to the latent cooling energy use. A common measure
of cooling capacity is cooling tons. There are 12,000 BTU in
one cooling ton, which is the heat energy absorbed by a ton of
ice with the melting occurring at constant temperature of 32 F.
Latent Cooling Energy Use (BTUh) = 1,076 × Q ×
.075 lbm/cubic ft. × (w2 – w1) × 60 min/hr
EXAMPLE PROBLEM 11-1 (Heat Energy Demand) (IP Units)
What is the heating energy demand to raise the temperature
of 10,000 SCFM supply air from 0 F to 120 F?
Heating Energy Use = 1.08 × Q × (T2 – T1)
= 1.08 × 10,000 SCFM × (120 – 0) F
= 1,296,000 BTUh
11.4.2 Heating Systems. With the availability of piped natural gas, many new heating systems are of the direct gas-fired
type instead of heated water flowing through a coil.
The energy used to accomplish the heating and/or cooling
of the supply air is related to the airflow and the size of the
temperature difference imposed on the airflow. For the heating
requirement the following applies:
Heating Energy Use (BTUh) = 0.24 × Q × .075 lbm/cubic ft.
× (T2 – T1) × 60 min/hr = 1.08 × Q × (T2 – T1)
where:
0.24 = the specific heat of standard air
Q = airflow, SCFM
T2 = heated air temperature
T1 = air temperature before heated
When heating supply air, the moisture (amount of water
vapor) in the air is not typically a concern. Heating of air does
not change the amount of moisture content in the air; this is
called sensible heating. When the moisture content of the air is
changed, latent heating (when moisture is added) or latent
Air handling units (AHUs) are usually categorized according to the source of heat: steam, hot water, indirect gas and oilfired units, and direct gas-fired units. Table 11-3 summarizes
the basic differences of typical industrial AHU heating
approaches. Each type of air heater has specific advantages
and limitations that must be understood by the designer when
making a selection. Each type must be capable of constant
operation. Variations occur within each type in their capability
of delivering a wide range of air temperatures, but they should
be able to control the discharge air temperature within a range
of 5 F [3 C]. Hot water and steam coil types are better able to
achieve a narrow temperature range of desired room conditions due to superior modulation ability and low heat control.
11.4.3 Steam Coil Heating. Steam heating was used in the
earliest air heaters applied to general industry as well as commercial and institutional buildings (Figure 11-7). When properly designed, selected, and installed, they are reliable and safe
but need a lot of maintenance. They require a reliable source
of clean steam at a dependable pressure. The principal disad-
Supply Air Systems
11-13
TABLE 11-3. Summary of Advantages and Limitations of Typical Industrial Heating Sources
vantage of steam units is the high maintenance effort to keep
traps, valves and condensate pumps operating properly. Other
disadvantages are potential damage from freezing or water
hammer in the coils, the complexity of controls when close
temperature limits must be maintained, higher installed cost,
and excessive piping.
•
Size the traps and return piping for the maximum condensate flow at minimum steam pressure plus a safety
factor.
•
Provide atmospheric vents to minimize the danger of a
vacuum in the coil that would keep condensate from
draining.
Freezing and water hammer are the result of poor equipment selection and installation and can be minimized through
careful design.
•
Never permit the condensate to be lifted by steam pressure.
•
Size the coil to provide the desired heat output at the
available steam pressure and flow.
•
Consider using a steam distributing coil with vertical
tubes.
The majority of freeze-up and water hammer problems
relate to the steam modulating type of unit that relies on throttling of the steam supply to achieve temperature control. When
throttling occurs, a vacuum will be created in the coil; unless
adequate venting is provided, condensate will not drain and
11-14
Industrial Ventilation
steam valve. When, and only when, this valve is closed, the
modulating steam valve on the pre-heat or first coil begins to
close. It is never allowed to close to the point where the air temperature leaving the coil, measured by the freezestat senses a
temperature below its setting, usually 40 F [4 C].
FIGURE 11-7. Single steam coil unit
can freeze rapidly under the influence of cold outdoor air. Most
freeze-ups occur when outdoor air is in the range of 20 F to 30
F [-7 C to -1 C] and the steam control valve is partially closed,
rather than when the outdoor air is a minimum temperature and
full steam supply is occurring (Figure 11-8).
Safety controls are often used to detect imminent danger
from freeze-up. This is normally done by a freezestat – an
extended bulb thermostat on the downstream side of the coil
connected into the control circuit to shut the unit down when
the temperature falls below a safe condition. An obvious disadvantage is that the plant air supply is reduced; if the building should be subjected to an appreciable negative pressure,
unit freeze-up may still occur due to cold air leakage through
the fresh air dampers.
Steam heating should not be used where temperature control
is critical. Temperature control with steam coils is accomplished
by operating a valve that allows steam to flow into the coil. The
steam condenses and the water drains away through a steam
trap. This type of control is basically modulation of the steam
coil, which does not provide good close temperature control. To
improve temperature control, use two control valves instead of
one. One valve is usually sized for about two-thirds of the
capacity, and the other valve is sized for one-third of the capacity. Through suitable control arrangements, both valves will provide 100% steam flow when fully opened and various combinations will provide a wide range of temperature control. Controls
are complex in this type of unit, and care must be taken to insure
that pressure drop through the two valve circuits is essentially
equal.
Multiple coil steam units (Figure 11-9) and bypass designs
(Figure 11-10) are available to improve the temperature control
range and help minimize freeze-up. With multiple coil units, the
first coil (preheat) is usually sized to raise the air temperature
from the design outdoor temperature to at least 40 F [4 C]. The
coil is operated with an on-off valve that will be fully open
whenever the outdoor temperature is below 40 F [4 C]. The second (reheat) coil is designed to raise the air temperature from 40
F [4 C] to the desired discharge condition. Refined temperature
control can be accomplished by using a second preheat coil to
split the preheat load. When less heat is required, it is best to
reduce steam flow to the second or reheat coil by a modulating
Bypass units incorporate dampers to direct the airflow.
When maximum temperature rise is required, all air is directed
through the coil. As the outdoor temperature rises, more and
more air is diverted through the bypass section until finally all
air is bypassed. The principal disadvantage of this type of unit
is the bypass is not always sized for full airflow at the same
pressure drop as through the coil, thus (depending on the
damper position) the unit may deliver differing airflow rates.
Damper airflow characteristics are also a factor. An additional
concern is that in some units, the air coming through the bypass
and entering the fan compartment may have a nonuniform temperature characteristic that might affect the ability to deliver air
within a close temperature range.
Another type of bypass design, called integral face and bypass
(Figure 11-11), features alternating sections of coil and bypass.
This design promotes more uniform mixing of the air stream,
minimizes any nonuniform flow effect, and, through carefully
engineered damper design, permits minimum temperature pickup of about 3 F [2 C], even at full steam flow and full bypass.
The same basic control system that has proven satisfactory for
a two-coil system can be used for a face and by-pass system.
The by-pass dampers are modulated closed when less heat is
desired. Then, and only then, is the steam flow reduced to the
coil by the steam modulating valve.
11.4.4 Hot Water Coil Heating. Hot water is an excellent
heating medium for air heaters. As with steam, there must be a
dependable source of water at predetermined temperatures for
accurate coil sizing. Hot water units require less maintenance
and are less susceptible to freezing than steam because the
pumped water flow ensures that the cooler water can be positively removed from the coil. Many large plants use hot water
systems that have temperatures above that of boiling. A pressure
is put on the pipe system to keep the water from flashing to
steam. Water temperatures are reduced at the heating units using
water to water heat exchangers. For a 100 F [38 C] air temperature rise and an allowable 100 F [38 C] water temperature
drop, 1 gpm [0.06 liters/s] of water will provide heat for only
450 acfm [0.23 am3/s] of air. This range can be extended with
high temperature hot water systems.
Temperature control for all applications is excellent with hot
water coils. Temperatures are easily maintained in a narrow
range since the temperature of the hot water can be varied. The
operation of the coil control valve to reduce or increase flow
for temperature changes does not need to be as precise as with
a steam coil.
Hybrid systems using an intermediate heat exchange fluid,
such as ethylene glycol and water mixtures, also have been
installed by industries with critical air supply problems and a
Supply Air Systems
11-15
11-16
Industrial Ventilation
FIGURE 11-9. Multiple coil steam unit
desire to eliminate all freeze-up dangers. A primary steam or
hot water system provides the necessary heat to a converter
that supplies a secondary closed loop of the selected heat
exchange fluid. The added equipment cost is at least partially
offset by the less complex control system.
11.4.5 Indirect Gas/Oil-fired Heating. Indirect gas/oil-fired
units (Figure 11-12) are widely applied in small industrial and
commercial applications. Indirect-fired heaters incorporate a
heat exchanger, commonly stainless steel, which effectively
separates the incoming air stream from the products of combustion. The gas/oil is burned inside the heat exchanger and
the supply air being warmed passes over the outside. Positive
venting of combustion products is usually accomplished with
induced draft fans. The indirect-fired air heater permits the use
of room air recirculation since the air stream is separated from
the products of combustion. This separation also allows oil to
be used as a heat source. Since the supply air is not exposed to
an open flame, this type of heater is well suited to ventilate
areas such as paint mix rooms and storage areas that have
potentially explosive fumes released in the workspace.
Temperature control, “turn-down ratio,” is limited to about
3:1 or 5:1 due to burner design limitations and the necessity to
maintain minimum temperatures in the heat exchanger and
FIGURE 11-10. By-pass steam system
flues. Turn-down ratio is a function of the heater’s ability to
modulate gas delivery from full gas delivery to zero (idle). If
the burner design and other features permit a 50% reduction of
gas delivery to the heater, the turn-down ratio is 2:1. If gas
delivery can be reduced to 25% of the maximum and the burner still operates satisfactorily, the turn-down ratio is 4:1. The
turn down ratio relates to the variability of raising the air temperature. So with a unit having a maximum temperature rise of
100 F [56 C], a 4:1 turn down would result in a 25 F [14 C]
rise minimum. Temperature control can be extended through
the use of a bypass system similar to that described for single
coil steam air heaters. Bypass units of this design offer the
same advantages and disadvantages as the steam bypass units.
Another type of indirect-fired unit is a condensing furnace.
With this type of heater the products of combustion are
dropped below the condensing temperature of water by cool
incoming supply air. The efficiency is often greater than 90%.
11.4.6 Direct Gas-fired Heaters. Direct-fired heaters, where
natural or liquid petroleum gas (LPG) gas is burned directly in
the air stream and the products of combustion are released in the
air supply, have been commercially available for some years
(Figure 11-6). Like the indirect gas/oil heating system the piping in the plant is much smaller compared to a steam or hot
water system. These units are economical to operate since all of
the heating value of the fuel is available to raise the temperature
of the air. This results in a net heating efficiency over 90+%.
Commercially available burner designs provide turn-down
ratios from approximately 25:1 to as high as 45:1 permitting
good temperature control.
In sizes above 10,000 acfm [5.00 am3/s], the units are relatively inexpensive on a cost per acfm basis; below this capacity, the costs of the additional combustion and safety controls
weigh heavily against this design. A further disadvantage is
that governmental codes prohibit the recirculation of room air
across the burner. Controls and sensors in these units are
designed to provide 1) a positive proof of airflow before the
burner can ignite, 2) a timed pre-ignition purge to insure that
any leakage gases will be removed from the housing, and 3) a
constantly supervised flame operation that includes both
flame controls and high temperature limits. For safety purposes, the flame controls have a number of pressure sensors and
valves in the gas piping to stop flow if significant changes in
gas pressure are experienced.
Concerns are often expressed with respect to potentially
toxic concentrations of carbon monoxide, oxides of nitrogen,
aldehydes, and other contaminants produced by combustion
and the resulting gases released into the supply air stream.
Practical field evaluations and detailed studies show that with
a properly operated, adequately maintained unit, carbon
monoxide concentrations should not exceed 5 ppm, and oxides
of nitrogen and aldehydes should be well within acceptable
limits.(11.2) Before specifying direct-fired equipment, evaluate
all the expected contaminants to determine if direct-fired heat-
Supply Air Systems
11-17
FIGURE 11-11. Integral face and by-pass coil(11.4)
ing is appropriate in the space. For example, direct-fired heating should not be used in heating/ventilating paint mix rooms
or fiberglass lay-up operations.
A variation of this unit, known as a bypass design, has
gained acceptance in larger plants where there is a desire to circulate large airflows at all times (Figure 11-13). The large airflow is needed for summer ventilation with outdoor air to
reduce hot plant temperatures. In the heating season, the outdoor air amount is reduced by recirculating plant air in the airhandling unit. In the bypass design, controls are arranged to
reduce the flow of outdoor air with a certain percentage flowing across the burner and the balance of the airflow provided
by the entry of room air into the fan compartment. In this way
the fan airflow rate remains constant and circulation in the
space is maintained. It is important to note that the bypass air
FIGURE 11-12. Indirect-fired unit
does not cross the burner; only 100% outdoor air is allowed to
pass through the combustion zone. Controls are arranged to
regulate outdoor airflow to insure that burner profile velocity
(the rate of airflow through the burner plates) remains within
the limits specified by the burner manufacturer — usually in
the range of 2,000 to 3,000 fpm [10 to 15 m/s]. This is accomplished by providing a variable profile that changes area as the
damper position changes. A similar type of unit has a fixed
amount of outside air passing over the burner. This is mixed
with return or unheated outside air. The total amount of outside
air is varied to provide adequate replacement air and to
achieve a building positive pressure. The air passing over the
burner is heated to higher temperatures for mixing with the
unheated air. A minimum of 20 percent of the total air must
pass over the burner to maintain suitable carbon dioxide levels.
Direct-fired heaters are not well suited for heating areas at outside doors unless they operate continuously since it takes two
FIGURE 11-13. Direct-fired by-pass unit
11-18
Industrial Ventilation
to three minutes before it can deliver warm air. This time period is required to purge the unit, have the safety devices in the
natural gas line check themselves, and open the gas valve.
The related disadvantage of the direct gas-fired system is the
requirement to use outside air. Since outside air is brought into
the building, it must also be exhausted. In the situation where
there is enough process exhaust to remove the outside air,
which is heated by the burner, no extra energy is used. If there
is an excessive amount of supply air over the process exhaust,
the excess air must be heated and then exhausted. Extra energy
is used in this case to heat more outdoor air that is required.
Inasmuch as there are advantages and disadvantages to both
direct-fired and indirect-fired replacement air heaters,(11.2) a
careful consideration of characteristics of each heater should
be made. A comparison of the heaters is given in Table 11-4.
11.4.7 Air Cooling Equipment. Since most industrial facilities have a process heat release, the supply air system is
required to reduce the effect of this heat for temperature control
in summer or, in some locations, all seasons. The ability to use
untempered outside air to obtain space cooling depends upon
the amount of heat release from equipment in the space and the
outside air temperature. If the supply air temperature needs to be
lowered, air-cooling is accomplished by means of a cooling coil
(mechanical cooling) or an evaporative cooling unit. A detailed
discussion regarding air-cooling can be found in Chapters 20,
22, and 23 of the ASHRAE Handbook.(11.3) Cooling is utilized
for process requirements and to provide summer heat relief.
To provide summer relief of hot space temperatures, a greater
amount of outside supply air may be needed than that required
for replacement air purposes. In this situation, the use of cooling
may be justified since a lower airflow is required compared to
using untempered outdoor air ventilation to achieve reasonable
space temperatures. The use of outside air for cooling is calculated on a temperature rise of 20 F [11 C] before being exhausted. If a cooling unit is used, the entering temperature is lower,
allowing a supply air temperature rise of 30 F to 40 F [16 C to
22 C]. Thus, with cooling, less airflow is needed.
TABLE 11-4. Comparison of Heater Advantages and Disadvantages
Advantages
Disadvantages
Direct-fired Unvented:
1. Good turn-down ratio—8:1 in small sizes; 25:1 in large sizes.
Better control; lower operating costs.
1. Products of combustion in heater air stream (some CO2, CO,
oxides of nitrogen, and water vapor present).
2. No vent stack, flue or chimney necessary. Can be located in
sidewalls of the building.
2. Higher first cost in small size units.
3. Higher efficiency (90+%). Lower operating costs. (Efficiency
based on available sensible heat.)
4. Can heat air over a wide temperature range.
5. Lower first cost in large size units.
3. May be limited in application by governmental regulations.
Consult local ordinances.
4. Extreme care must be exercised to prevent minute quantities of
chlorinated or other hydrocarbons from entering air intake or toxic
products may be produced in heated air.
5. Can be used only with natural gas or LPG.
6. Burner must be of proven design tested to ensure low CO and
oxides of nitrogen content in air stream.
7. Outside air brought into building may be significantly more than
process exhaust causing an excessive amount of heating energy
use.
Indirect Exchanger:
1. No products of combustion are discharged into building.
1. Higher first cost in large size units.
2. Allowable in all types of applications and buildings if provided with
proper safety controls.
2. Turn-down ratio is limited — 3:1 usual, maximum 5:1.
3. Small quantities of chlorinated hydrocarbons will not normally
break down on exchanger to form toxic products in heated air.
3. Flue or chimney required. Can be located only where flue or
chimney is available.
4. Lower efficiency (80%). Higher operating cost.
4. Can be used with oil, LPG, and natural gas as fuel.
5. Can heat air over a limited range of temperatures.
5. Lower first cost in small size units.
6. Heat exchanger may be subject to severe corrosion condition.
Needs to be checked periodically for leaks after a period of use.
6. Can be used in air recirculation mode as well as for makeup air.
7. Difficult to provide combustion air from outdoors unless roof or
outdoor mounted.
Supply Air Systems
11-19
11.4.8 Mechanical Cooling. With mechanical cooling, the
cooling coil has a chilled fluid flowing through it to remove
the heat from the air stream. This heat exchange reduces the
temperature of the air stream and warms the chilled fluid. The
fluid is typically a refrigerant or water. Air handling units that
use a refrigerant have a compressor and condenser nearby to
change the refrigerant gas back into a liquid and reduce its
temperature. With this system, the act of quickly reducing the
pressure on the liquid allows it to change into a gas and
become cold, thus chilling the coil. In a chilled water unit,
water of approximately 45 F [7 C] flows through the cooling
coil. The water is chilled by a central chiller and pumped
through a pipe distribution system to each air-handling unit.
Commercial and light industrial type AHUs most often use the
refrigerant type system commonly called direct expansion
(DX) cooling equipment. The first cost of the chilled water
system is higher than the DX system, but it offers longer component life, reduced maintenance, lower energy costs, and is
more suitable for larger installations.
11.4.9 Evaporative Cooling. Evaporative cooling systems
rely upon the evaporation of water vapor into the air stream to
lower the air temperature. This also causes the air stream to
become more humid. In the evaporative cooling unit, air
absorbs water vapor as it passes through a wetted pad or
through a water spray zone. Energy is given up by the air to
evaporate the water and the air temperature is reduced. Since
evaporative coolers raise the relative humidity in the space,
this impact on the industrial processes should be evaluated.
Some evaporative cooling systems have their own pumps and
water circulating systems. Others rely on the pressure in the
water line to generate a water spray. Evaporative coolers are
commonly used in dry areas of the world but can be applied to
almost all areas of the United States. They are also used in
industrial applications that have high replacement airflow or
large internal heat releases. For an evaporative cooling unit to
operate at peak efficiency, the pads must be well wetted and
reasonably clean. Spray nozzles must be kept free of clogging
deposits. Equation 11.1 can be used to identify the temperature
leaving an evaporative cooler:
The use of a cooling coil can often reduce both air temperature and humidity. The humidity reduction is caused by dropping the air temperature below its dew point. The objective is
to get the air temperature cold enough so that the amount of
water vapor in the air can no longer be maintained. The air
begins to fog and water droplets called condensate begin to
form on the cooling coil. This condensing of water vapor to
reduce humidity requires additional cooling over and above
that for reducing the air temperature.
Texit = Tenter – E(Tenter – Tw)
EXAMPLE PROBLEM 11-2 (Cooling Energy Demand)
(IP Units)
What is the cooling energy demand to lower the temperature
of 10,000 SCFM supply air from 90 F to 60 F?
The 90 F air has a relative humidity of 50% with a moisture
content of 0.0153 lbm moisture per pound dry air. The
moisture content of saturated 60 F air is 0.0112 lbm moisture
per pound dry air.
Sensible Cooling Energy Use = 1.08 × Q × (T2 – T1)
= 1.08 × 10,000 SCFM × (90 – 60) F = 324,000 BTUh
Latent Cooling Energy Use = 1,076 × Q × .075 lbm/cubic ft
× (w2 – w1) × 60 min/hr
= 1076 × 10,000 × 0.075 × (0.0153 – 0.0112)
= 3,309 BTUh
Cooling Energy Use = Sensible Cooling Energy Use +
Latent Cooling Energy Use
= 324,000 BTUh + 3,309 BTUh = 327,309 BTUh
[11.1]
where
Tenter = Dry-Bulb temperature entering, F [C]
Texit = Dry-Bulb temperature leaving, F [C]
Tw = Wet-Bulb temperature entering, F [C]
E = Efficiency factor
The Wet-Bulb temperature is the value measured using a
psychrometer as discussed in Chapter 10. The efficiency is
normally 80%.
11.4.10 Air Filtration. Supply air filtration for workspaces
is not a major concern for most industrial processes; however,
seasonal factors such as insects, pollen, organic debris, etc.,
may require removal before the air is supplied. The filters are
typically selected on the basis of keeping the supply air unit
clean. However, in some cases, filters are selected for
employee health considerations or process concerns. Filters
for normal service typically have a minimum efficiency
reporting value (MERV) of 6 to 8 as defined in ASHRAE
Standard 52.2, “Method of Testing General Ventilation AirCleaning Devices for Removal Efficiency by Particle
Size.”(11.4) When a process requires a high level of cleanliness, such as food processing, painting, or assembly of parts
where a fine dust is a detriment, a more efficient filtration system is required. Refer to Chapter 8 for more discussion
regarding air cleaning equipment.
11.4.11 System Temperature Control. Some processes require a space that has close control of temperature and humidity. This often requires both heating and cooling of the supply
air to achieve the desired thermal conditions. These spaces
require humidification if the air is too dry, a condition that
most likely occurs in the winter. An example could be a pow-
11-20
Industrial Ventilation
der painting operation that requires air entering the paint spray
booth to be 70 F [21 C] and 50% relative humidity (RH). This
temperature and humidity condition is necessary to achieve
proper drying of the paint and to prevent arcing and sparks
inside the booth for fire prevention.
In summer, the supply air would need to be cooled below
51 F [11 C] to condense enough water vapor from the air to
achieve the 50% RH. This air would then need to be reheated
to raise the temperature to the 70 F [21 C] goal.
In winter, the air must be heated and water vapor added to
the air to achieve the desired 70 F [21 C], 50% RH. A heating
coil or gas-fired device can be utilized. Either a humidifier or
an evaporative cooler is used to add humidity. If a humidifier
is used, the heat in the water vapor must be identified and the
energy of the heater reduced accordingly. Figure 11-14 has a
representation of the performance of this equipment during the
cooling and heating seasons. For example, on a cold day the
air must be heated to a temperature of 99 F [37 C] to achieve
a condition of 70 F [21 C] and 50% RH.
The closeness of control desired will dictate the component
type to be utilized in the system. Heating and cooling water
coils provide the best control.
11.4.12 Unit Location. Air supply units are normally located in the upper level of the plant or on the roof. In some
designs, these units are placed just below the roof (in the
truss space) and have a catwalk system for ease of access.
Rooftop units create the need for people to walk on the roofs.
It is good practice to provide a walkway to minimize excessive wear on the single-ply roofs in common use today. Some
systems have the unit placed along an outer wall inside the
FIGURE 11-14. Air heating and cooling requirements
building. Outside air is mixed with room air to satisfy general building heating and replacement air requirements.
11.4.13 Size and Cost Considerations. There are several
cost considerations to a supply air system installed in an industrial facility. First, the relative cost for the supply air unit
decreases as the size increases. Some cost elements of the unit
increase with unit size: the unit housing, fan, filters, and coils.
The unit’s control cost depends on the control functions being
performed and is approximately the same for all size units.
Another major cost element is the air distribution system; i.e.,
the duct and registers. The duct and register costs increase as
the system gets larger. The final cost consideration is installation, which includes lifting the unit; structural steel supports,
electrical, natural gas, and other piping system hook-ups; unit
start-up; and warranty. Installation cost is somewhat independent of unit size and increases at a rate slower than the unit size.
For more information regarding system costs, see Chapter 2,
Cost Estimating.
11.5
SUPPLY AIR DISTRIBUTION
In an industrial facility, the supply air distribution plays an
important role in the success of controlling airborne contaminants. If contaminants are controlled by local exhaust ventilation, the supply/replacement air should be introduced into the
space in a way that does not interfere with the capture effectiveness of the exhaust hoods. Interference is created when
supply/replacement air is introduced at an excessive velocity
into the vicinity of an exhaust hood, thus interrupting the protective flow path of the hood’s exhaust air volume. When the
supply/replacement air diffuser is located too close to the
Supply Air Systems
exhaust outlet, the clean air may be “short-circuited” and not
reach the workspace at all.
There are additional supply air design considerations when
dilution ventilation is used rather than local exhaust ventilation
to control contaminants. These include the location of the supply air outlets, the rate of airflow, and the placement of the
exhaust air intakes. Refer to Chapter 10 for more discussion and
system sizing considerations. The choice of dilution ventilation
versus local exhaust ventilation depends on the nature and
quantity of the contaminants and the workspace. Several supply
air design approaches are discussed in the following sections.
11.5.1 Unidirectional or Plug Airflow. The use of non-turbulent or laminar supply airflow is required in situations
where high cleanliness or extreme contaminant control is
desired. This approach has clean supply air moving across the
space in a uniform direction and the air is removed from the
space at a location opposite the supply air entry point. This
design scheme is often referred to as unidirectional, laminar, or
plug airflow. It is normally employed to protect workers and
critical processes. In addition to careful consideration of the
supply air distribution design, physical obstructions such as
partitions or furniture should be minimized to avoid any turbulent airflow. Examples of this type of supply air design can be
found in industries or activities associated with firing ranges,
pharmaceutical manufacturing, semiconductor manufacturing,
healthcare treatment, aerospace, and painting operations.
For areas that require non-turbulent air for proper exhaust
system operation, one approach is to pass air through a supply
air plenum built as part of a perforated ceiling and/or through
perforated duct. The ceiling plenum or duct runs should cover
as large an area as possible to diffuse the airflow. A plenum
wall providing cross-flow ventilation should be used when the
workers are positioned between the supply air system and the
contaminant source or exhaust hood. This approach should not
be used for design velocities at the worker over 100 fpm [0.5
m/s] since a low pressure zone can be created causing contaminants to be carried into the worker’s breathing zone. See
Chapter 6, Section 6.1.4 for more information on Worker
Position Effects.
11-21
lence problems similar to large diffusers. This can cause reentrainment of contaminants from the room into the clean
replacement air via a low-pressure area created near the introduction point. The low-pressure phenomenon also creates
uneven replacement air distribution in the room. Providing a
wide replacement air plenum and slowly introducing supply air
into the plenum will reduce the problem. However, space for a
wide plenum is frequently unavailable. One solution is to feed
the plenum with a perforated duct to diffuse the air inside the
plenum. Ensure that the proper pressure adjusting devices (e.g.,
orifice plates) are installed per the manufacturer’s recommendations. Another approach is to distribute air from either a ceiling or wall-mounted plenum and design the plenum face with
two overlapping perforated plates, one fixed and one
adjustable, airflow for balancing, located 2 to 6 inches [50 to
150 mm] apart. Air flowing through slightly offset holes will
encounter more resistance; thus, air quantities passing through
the low-flow areas will increase. The holes must be small
enough to fine-tune the airflow from the plenum. Openings of
3/8" [9 mm] diameter in the adjustable plates with sufficient
numbers to provide a velocity of 2000 fpm [10 m/s] seem to
work well.
The second approach is used in clean room and paint booth
designs to achieve a high control on air cleanliness. For these
applications, clean supply air flows through a grid of filters in
the ceiling and is exhausted at floor level. Flow velocities in
the range of 50 to 100 fpm [0.25 to 0.50 m/s] are common.
11.5.2 Mixing Ventilation Systems. The mixing approach
to the supply air ventilation system relies on high-velocity air
streams leaving supply registers as the means of delivering air
to the workspace. These jets of supply air quickly entrain and
mix with the space air. As shown in Figure 11-15, the average
temperature of this air stream begins to approach the space
temperature as the velocity of the jets slows. This example has
air leaving a register at a velocity of 2000 fpm [10 m/s] and a
Perforated drop-type ceilings work best in spaces with ceiling heights of less than 15 feet [4.5 m]. Hoist tracks, lighting,
and fire protection systems can be built into the ceiling. In
some cases, fire protection will be required above and below
the ceiling. Use the perforated duct approach when ceiling
heights are over 15 feet [4.5 m]. Perforated duct manufacturers
typically have computer programs to assist designers in determining duct sizes, shapes, and types as well as the location of
pressure adjusting devices such as orifice plates and reducers.
Airflow delivery in large bays may require supplemental air
delivered at workstations to provide comfortable conditions for
workers.
How the supply air is fed into a plenum is critical to its performance. High velocity flow into the plenum can cause turbu-
FIGURE 11-15. Air jet temperature and velocity profile
(IP units)
11-22
Industrial Ventilation
temperature of 20 F [11 C] below room temperature. At a
workstation 27 feet [8 m] from the register, the average speed
of the jet has dropped to 200 fpm [1.0 m/s] and the air temperature has approached the room temperature of 86 F [30 C].
The actual conditions depend on the register selected, but
velocities of 100 to 200 fpm [0.5 to 1.0 m/s] and temperatures
of one to two degrees below the room temperature are likely.
Mixing systems dilute airborne contaminants the same way
that the air jets dilute temperature. Care should be taken to
direct the supply air jets so as not to disturb the performance
of local exhaust systems. Otherwise the resulting air currents
can sweep contaminants away from exhaust hoods rendering
the hoods less effective. The local exhaust hoods may then
require additional airflow to control the contaminants.
Increasing exhaust airflow also increases energy costs due to
the need for larger fans and motors. In extreme cases of high
room air motion, the hoods remain ineffective even with substantial increases in airflow. Hence, workers could still be
overexposed even with a local exhaust system in place.
Therefore, supply air discharge should be located away from
local exhaust hoods.
to the workstation. Even better is the direction of the supply air
up through the workstation from a grate on the floor. Both of
these approaches have the air discharge much closer to the
worker so little entrainment of room air takes place.
Care must be taken with spot cooling systems. Air delivery
at high velocities from behind the operator will create a low
pressure zone in front of the body (the person’s breathing
zone). Contaminants can be induced into this zone and inhaled
by the worker. Care must be taken also not to blow contaminants into the employee’s eyes, so spot cooling should not be
used with processes that have airborne particles escaping. Spot
cooling systems for these applications often have airflows in
the range of 3,000 to 4,000 acfm [1.5 to 2.0 am3/s] per workstation and velocities at 1500 to 2000 fpm [7.5 to 10.0 m/s]
leaving the supply air register.
Another common problem with airflow in an industrial setting relates to the inappropriate use of pedestal fans. These
spot cooling fans are often used to provide cooling air movement directly at an employee position. However, unless they
are carefully directed, they may impact the effectiveness of
local exhaust hoods.
11.5.3 Air Displacement Ventilation Systems. A nonturbulent approach to adding air into the workspace is called
air displacement. Air displacement ventilation systems were
first applied in the welding industry in 1978, and now are
widely used in Scandinavian countries. This type of supply
air system relies on process heat to warm the air and make
it rise. Provision is made to remove the warm air near the
ceiling of the space. The supply air is introduced into the
space through low-velocity diffusers placed near the floor.
The objective of the air displacement system is to achieve
air quality conditions in the occupied zone that are similar to
those of the supply air.
Mixing systems can have air outlets in the truss space (20
feet [6 m] or higher) blowing downward or placed at lower
levels. For those systems where the air is discharged below the
truss, duct routing must be coordinated with the process layout
and the needs of the process equipment. Quite often the use of
cranes, gantries, conveyors, and other material handling equipment greatly reduces the access to space for routing duct
below the truss. A common low-level discharge height is 10
feet [3 m] above the floor with the air directed horizontally.
During the summer there is a downward deflection of the supply jet, while in the winter, the air is directed upward approximately 5 degrees above horizontal. The lower-height-discharge approach provides a cooler workspace and should be
considered when a lower space temperature is desired.
As illustrated in Figure 11-16, there are two air distribution
zones in an air displacement system, the upper and lower strata.
The upper zone is formed at the elevation where the supply air
quantity equals the total air moving upward in the thermal
plumes caused by the process heat. As this warm air rises, it
entrains adjacent air and the total volume of moving air increases. When this total air volume equals the supply air, there is no
more incoming air to feed the plume and recirculation of space
air begins. The elevation where the recirculation starts is called
the stratification level. Properly designed air displacement systems have the stratification level well above the occupied lower
zone. The height of this lower zone is dependent on the amount
of supply air, the nature of the heat sources, and the air distribution across the floor.
Supply air not removed by process exhaust systems is normally removed from the building through the use of roofmounted exhaust fans. The discharge of supply air at the 11foot [3 m] level is an approach often used for spot cooling.
Spot cooling is the directing of a mass of supply air to a workstation with the purpose of keeping it as cool as possible. It is
often used in operations that have high radiant heat exposures
such as is found in metal casting, forging and steel making
operations. The approach of high velocity discharge spot cooling is normally not very effective if the air discharge grille is
located a distance from the workstation. More effective methods place the supply air outlet at the same height and adjacent
When designing a displacement ventilation system, the following parameters need to be considered: 1) supply airflow rate
and temperature; 2) air temperature at floor level; 3) vertical
temperature gradient; 4) maximum air velocity at floor level;
and 5) first cost, operating cost, and energy consumption.(11.5)
The supply air temperature can be 4 F to 6 F [2 C to 3 C]
warmer than that used in a mixing type system to achieve the
same occupied space temperature.(11.6) The vertical temperature gradient or the temperature rise of the supply air compared to the exhaust is greater in the displacement type system. Typical temperature differences compared with increases
in building height are:(11.7)
Supply Air Systems
11-23
FIGURE 11-16. Airflow in displacement ventilation system
Building Height, ft [m]
Less than 10 [3]
10–20 [3–6]
over 30 [> 9]
Temperature Rise, F [C]
11–13 [6–7]
15–18 [8–10]
18–22 [10–12]
This increase in temperature difference will reduce the amount
of exhaust air required.
The advantage of not significantly mixing the space air with
the supply air is a workspace that is cooler and has less airborne contaminants. The process heat and many of its associated contaminants are carried away as the warm air rises.
Special provisions must be made for supply air outlets. Since
they are on the floor, they must be coordinated with the process equipment layout to allow access to operate, service and
maintain the equipment. Air outlets need to be placed a reasonable distance from each other to avoid drafts caused by the
high quantity of supply air leaving the diffusers.
11.5.4 Duct Materials. Supply duct materials are generally
Sheet Metal and Air Conditioning Contractors National
Association (SMACNA) Class I or II medium gauge sheet
metal, but other materials such as specially coated cloth, may
be used. The material does not need to be as strong as exhaust
duct for several reasons:
1) It is not exposed to the transport of abrasive process
contaminants.
2) The system operates at a relatively low pressure.
3) Much of the duct is on the downstream side of the fan
and is under a positive pressure.
4) Duct leaks do not pose a health hazard and have little
effect on system performance when placed inside the
building.
The duct needs to be strong enough to last in its environment. Often the abuse of plant operations requires a heavier
duct system than one that is hidden above a ceiling.
If the sheet metal gauge selected is too light, fan noise may
be more pronounced due to vibration of the duct. Metal stiffeners attached to the outside of the duct help prevent this type
of noise. SMACNA standards provide detailed information on
duct construction.(11.8, 11.9)
11.5.5 Sheet Metal. Sheet metal materials typically include
galvanized steel, uncoated steel, stainless steel and aluminum.
In addition to round ducts, oval or rectangular ducts are often
used in order to adjust the cross-sectional area to avoid
obstructions in the building space. SMACNA standards use
pressure ranges to classify duct thickness and construction
methods. The allowable duct leakage and acoustical considerations also dictate the construction methods.
11.5.6 Plastic. Thermosetting (glass-fiber reinforced polyester) and thermoplastic (polyvinyl chloride, polyethylene,
polypropylene, acrylonitrile butadiene styrene) construction
standards are provided by SMACNA’s Thermoplastic Duct
Construction Manual.(11.10)
11.5.7 Fiberglass. For acoustical reasons, supply duct may
be lined for better sound absorption. The lining used is a fiberglass-based material treated to minimize moisture absorption.
The coating is also mold and fungus resistant.
11.5.8 Textile. For certain applications, especially those
with high open bays and/or areas requiring frequent cleaning,
porous textile duct is a relatively new option. The supply air
inflates the ductwork. The textile duct can be washed and
reused and ultimately recycled. Hanging systems rely on engineered orifices, varying fabric porosity and linear vents to distribute airflow.
11-24
Industrial Ventilation
11.5.9 Supply Air System Design Considerations. A
properly designed ventilation system must adhere to building
codes, requirements established by the National Fire Protection
Association (NFPA), and standards developed by SMACNA.
Those related to energy use should be thoroughly reviewed
since they can impose restrictions on the ventilation system
design. There are also standards published by ASHRAE,
ANSI, and AMCA as well as those developed by specific
industrial corporations.
There are several design tools currently available to aid in
the design of industrial ventilation systems. These design tools
are used to evaluate building air balance, heating and cooling
loads, special pressures, smoke/contaminant migration, and
migration and dilution of gases and/or fumes.
When there is a need to understand ventilation system performance for contaminant control, designers sometimes use
computational fluid dynamics (CFD). This approach allows
multiple scenarios to be analyzed with limited investment.
With CFD, the geometry of a space is configured and airflow
rates, temperatures, contaminant migration, and exhaust capture efficiency can be evaluated. The CFD model divides time
and distance into discrete intervals. The modeled space is
divided into many smaller volumes that interact with each
other based on fundamental conservation equations. An iterative process is used to calculate the results and computation
times can take a number of hours depending on the complexity
of the problem and the capacity of the computer.(11.11)
There are numerous computer programs, nomagraphs, and
other catalog-type data available from manufacturers for selecting and sizing system components. Air handling unit fans,
coils, filters, and humidifiers are selected with this information.
Elements in the air distribution system (duct, registers, etc.) are
also selected with these aids. It should be noted that supply air
duct pressure loss calculations are much easier to accomplish
than those for exhaust air systems. The air is clean and dampers
are often used to adjust airflow through the duct system. Duct
velocity is typically limited to 2500 to 3000 fpm [12.5 to 15.0
m/s] to avoid excessive noise and minimize horsepower
requirements of the fan motor. Registers are selected to obtain
the desired air throw or air distribution in the space.
Industrial rated duct components should be used in air distribution systems since the higher velocities being utilized will stress
the materials found in lesser duty components. Balancing
dampers may not be required, but if used in a system that has air
velocities greater than 2000 feet per minute [11.0 m/s], industrial
rated dampers are needed. For systems that serve large industrial
spaces, often an air balance is not critical to the facilities operation,
and in these situations, dampers provide little value to system performance. Dampers should not be installed behind grilles to minimize noise and stress on the blades found in a grille or diffuser.
11.6
AIRFLOW RATE
The design supply airflow rate depends on several factors,
including health and comfort requirements. Sensible heat can
be removed through simple air dilution. Nuisance or undesirable contaminants can also be reduced by dilution with outdoor air. These topics are described in Chapter 10. For many
industrial facilities, experience shows that when the air supply
is properly distributed to the working level (i.e., in the lower 8
to 10 ft of the space [2.4 to 3.0 m]), outdoor air supply of 1 to
2
2 acfm/ft [0.005 to 0.01 am3/s/m2] of floor space will give
good results. This flow rate will normally satisfy the process
exhaust quantity as well as circulate adequate air for building
heating requirements and general ventilation. Specific quantities of minimum outdoor air supply may be obtained from
building and health codes or from criteria developed by groups
such as ASHRAE.
11.6.1 Air Changes. The number of air changes per minute
or per hour is the ratio of the airflow ventilation rate (per
minute or per hour) to the room volume. This normally applies
to the flow of outdoor air. Air changes per hour or air changes
per minute is a poor basis for ventilation criteria. The required
ventilation depends on the generation rate and toxicity of the
contaminant, not on the size of the room in which it occurs. For
example, let us assume a situation where an airflow of 11,650
acfm [5.497 am3/s] would be required to control solvent vapors
by dilution. The operation may be conducted in either of two
rooms, but in either case, 11,650 acfm [5.497 am3/s] is the
required ventilation. The air changes, however, would be quite
different for the two rooms. As can be seen in Table 11-5, for
the same air change rate, a high ceiling space will require more
ventilation than a low ceiling space of the same floor area.
Thus, there is little relationship between “air changes” and the
required contaminant control. Also, there are often great differences between the calculated air change per hour and the effective air changes, which is determined by location of air supply
outlets, exhaust systems, crossdrafts and air temperature.
Room air change rate is the same as dilution ventilation that is
discussed in Chapter 10. The difference between the actual and
effective air change rate is the same as applying mixing factor
(mi) values to increase the amount of airflow because of less
TABLE 11-5 (IP). Air Exchanges vs. Room Size
Room Size
Room ft3
Air changes/
minute
Air changes/
hour
40 H 40 H 12 high
19,200
11,650/19,200 = 0.61
36
40 H 40 H 20 high
32,000 11,650/32,000 = 0.364
22
TABLE 11-5 (SI). Air Exchanges vs. Room Size
Room Size (m)
Room (m3)
Air changes/s
12.192 H 12.192 H 3.657
543.5943
5.479/543.5943 = 0.0101
12.192 H 12.192 H 6.096
906.1391
5.497/906.1391 = 0.0061
Supply Air Systems
than ideal ventilation conditions.
The air change basis for ventilation does have applicability
for relatively standard situations such as office buildings and
school rooms where a standard ventilation rate is reasonable.
Building codes for the design of certain types of buildings
often use a minimum air change per hour for specific spaces.
For example, a flammable storage room requires six air
changes per hour using OSHA requirements. This approach is
easily understood and reduces the engineering effort required
to establish a design criteria for ventilation. It provides a minimum ventilation rate and is acceptable when other information
that would lead to greater ventilation rates is lacking.
11.7
HEATING, COOLING AND OTHER OPERATING
COSTS
Operating the supply air system can be a major expense for
a manufacturing plant. In addition to heating the air during the
cold weather and possibly cooling in the hotter months, there
are other operational concerns. Perhaps the highest energy
user is the electric motor that creates the air movement. A
rough estimate for motor size is one horsepower for every
FIGURE 11-17. Register airflow patterns(11.13)
11-25
1,000 acfm of airflow [approx. 1500 watts/am3/s]. For a complete discussion of ventilation system energy, see Chapter 12.
Other operational costs are air cleaning component replacement, control calibration, and maintenance.
11.7.1 Estimating Heating Energy Use. Supply air tem-
perature is controlled by the demand for heating and cooling.
These are the factors to consider in maintaining a comfortable
work environment for occupants: setpoint temperature,
humidity control, air distribution, and airflow rate. Where high
internal heat loads are to be controlled, a low air supply temperature can be obtained by reducing the amount of heat supplied to the air during the winter months and by deliberately
cooling the air in the summer. When a large airflow rate is
delivered at or below space temperatures, air distribution is
very important to maintain satisfactory conditions for individuals.
Maximum utilization of the supply air is achieved when the
air is distributed in the living zone of the space below the
10 foot [3 m] level (Figure 11-17).(11.3) When delivered where
the majority of people and processes are located, maximum
ventilation results with minimum airflow. It is important to con-
11-26
Industrial Ventilation
sider the comfort of the occupants when delivering air to the
space. In hot workspaces that use outdoor air for ventilation,
higher velocities may be acceptable in the summer (see Chapter
10). If the air is cool, however, the high velocity can cause discomfort. For hot workspace ventilation, the air can be distributed uniformly in the space or where required for worker comfort.
Heavy-duty, adjustable, directional grilles and louvers are desirable in this situation since they allow individual workers to
direct the air as needed.(11.1) Light gauge, stamped grilles
intended for commercial use are not satisfactory. Suitable control must be provided to accommodate seasonal and even daily
requirements. Warning: When the work process generates a
hazardous contaminant, worker adjustable supply grilles may
be incompatible with the contaminant control scheme.
11.7.2 Air Supply vs. Plant Heating Costs. Even if the supply air was drawn into the building through openings by the
action of the exhaust fans, heat would still be required at worker
stations and locations. Normally such a system would use unit
heaters for building temperature control and air would be supplied by numerous openings in the building outside. However,
experience has shown that introducing the same quantity of outdoor air through properly designed supply replacement air
heaters results in the same or lower fuel cost than using unit
heaters. The unit heater is typically located along the inside
perimeter of the building. With cold air coming into the building
through openings, the unit heater runs an excessive amount of
time, and the building as a whole is overheated.
11.7.3 Energy Considerations. The cost of heating supply air is a significant portion of the annual operating cost of
a ventilation system. Processes requiring cooling also need to
be evaluated for their energy and operating costs. However,
occupant comfort is more important than saving a few dollars
in energy costs. Recent indoor air quality studies quantify
diminished productivity when workers are uncomfortable. A
compilation of nine productivity and temperature reports
indicates that people are most comfortable between 72 F and
77 F [22 C and 25 C].(11.12) Acclimated industrial personnel
may tolerate slightly higher temperatures. In addition to the
equipment first cost, local building codes, and environmental
regulations, designer experience in utility incentives and operating costs are involved in purchasing decisions.
Cooling Energy Considerations. When heating, the only
energy considered is that required to raise the air temperature.
That is the sensible heat energy. If the quantity of water vapor
in the air is changed, a latent energy change also takes place.
In most cooling system operations, a latent energy change
takes place to lower the air humidity level. This latent energy
change occurs when the air is cooled to a temperature below
the dew point of the air. When this occurs, water is condensed
from the air stream onto the cooling coil.
In plant operations, the humid air can come from the outside
or be created by a process that releases water vapor into the
general air.
The cooling system is sized to overcome several energy
sources: the heat released to the space by the process (this
varies with production rate), the heat absorbed by the building
structure (since the outdoor air is warmer than the inside space
temperature), the heat absorbed by the building structure from
solar radiation, and the heat caused by the outside supply air
being warmer than the desired space temperature.
Filter Replacement. As the filter works to remove airborne
material from the air stream, it becomes saturated with that
material, and the pressure drop for air to pass through the filter
increases. As the pressure drop rises, the airflow decreases
unless the fan speed is increased or other adjustments are
made. Unless a replacement filter is installed, the original filter
will eventually become blinded allowing little or no air to pass.
Refer to Chapter 8 for more information regarding supply air
filter efficiency and application by type.
11.7.4 System Maintenance. Maintenance of the supply
air system is mainly associated with the supply air unit and
its operating controls. In addition to filter replacement, there
are fan bearings to grease and belts to replace. Coils and
drain pans must be cleaned and dampers lubricated. If the
system includes humidifying equipment, there are spray nozzles to clean/replace; there may be a pump to service as well
as controls to check and adjust. A major maintenance issue is
the recalibration of the systems controls. If the controls are
not routinely calibrated, operating set points may drift,
resulting in energy waste and poor system performance.
11.7.5 Untempered Air Supply. In many industries utilizing hot processes, cold outdoor air is supplied untempered or
moderately tempered to dissipate sensible heat loads from
the process and to provide temperature relief for workers.
The air required for large compressors, as well as for cooling
tunnels in foundries, can also come directly from outside the
plant and thus eliminate a heating load that would occur if
tempered replacement air was used.
11.7.6 Energy Recovery. Energy recovery from exhaust
air is accomplished through the use of heat exchange equipment to extract heat from the air stream before it is exhausted
to the outside. The application of return or recirculated cleaned
air from industrial exhaust systems is another method of
recovering process heat for use in the building. The application of heat exchangers to industrial exhaust systems is discussed in Chapter 12.
11.8
INDUSTRIAL EXHAUST RECIRCULATION
Where large amounts of air are exhausted from a room or
building in order to remove particulate, gases, fumes, or
vapors, an equivalent amount of fresh tempered replacement
air must be supplied to the room. If the amount of replacement
air is large, the cost of energy to condition the air can be very
high. Recirculation of the exhaust air after thorough cleaning
is one method of reducing the amount of energy consumed.
Acceptance of such recirculating systems will depend on the
Supply Air Systems
degree of health hazard associated with the particular contaminant being exhausted as well as other safety, technical, and
economic factors. A logic diagram listing the factors that must
be evaluated is provided in Figure 11-18.(11.14)
Essentially this diagram states that recirculation may be permitted if the following conditions are met:
1) The chemical, physical, and toxicological characteristics of the chemical agents in the air stream to be recirculated must be identified and evaluated. Exhaust air
11-27
containing chemical agents whose toxicity is unknown
or for which there is no established safe exposure level
should not be recirculated. Exhaust air from processes
using or generating explosive agents should never be
recirculated.
2) All governmental regulations and relevant standards
regarding recirculation must be reviewed to determine
whether recirculation is restricted or prohibited for the
system under review.
3) The effect of a recirculation system malfunction must
be considered. Recirculation should not be attempted if
a malfunction could result in exposure levels above
published OELs. Substances that can cause permanent
damage or significant physiological harm from a short
overexposure should not be recirculated.
4) An air cleaning device capable of providing an effluent
air stream contaminant concentration sufficiently low
to achieve acceptable workplace concentrations must
be available. Also, it must not impact work conditions.
For example, a scrubber or mist collector may provide
suitable cleaning efficiency but adds significant humidity to the air.
5) The effects of minor contaminants should be reviewed.
For example, welding fumes can be effectively
removed from an air stream with a fabric filter; however, if the welding process produces oxides of nitrogen,
recirculation could cause a concentration of these gases
to reach an unacceptable level.
6) Recirculation systems must incorporate a monitoring
system that provides an accurate warning or signal
capable of initiating corrective action or process shutdown before harmful concentrations of the recirculated
agents build up in the workplace. Monitoring may be
accomplished by a number of methods and must be
determined by the type and hazard of the substance.
Refer to Chapter 6, “Monitoring and Maintenance –
Air Cleaning Devices” in the O&M Manual. Examples
include area monitoring for nuisance type substances
and secondary high efficiency filter pressure drop indicators or on-line monitors for more hazardous materials.
11.8.1 Evaluation of Employee Exposure Levels. Under
equilibrium conditions, the following equations may be used
to determine the concentration of a contaminant permitted to
be recirculated in the return air stream:
CR = [(1 – h)(CE – KRCM)] / [1 – (KR)(1 – h)]
FIGURE 11-18. Recirculation decision logic(11.14)
where: CR = air cleaner discharge concentration after
recirculation
h = fractional air cleaner efficiency
CE = local exhaust dust concentration before
recirculation
[11.2]
11-28
Industrial Ventilation
KR = coefficient which represents a fraction of the
recirculated exhaust stream that is composed
of the recirculated air returning from the air
cleaner (range 0–1.0)
CM = replacement air concentration
NOTE: Units for CR, CE and CM are in parts per
million (ppm) or milligrams per cubic meter (mg/m3)
and all must be in the same unit system.
CB = [(QB/QA) (CG – CM)(1 – f)] + [ (CO – CM) f ]
+ [KBCR + (1 – KB)(CM)]
[11.3]
where: CB = 8-hr TWA worker breathing zone
concentration after recirculation
QB = total ventilation airflow before recirculation
(acfm) [am3/s]
QA = total ventilation airflow after recirculation
(acfm) [am3/s]
CG = general room concentration before
recirculation
f = coefficient which represents the fraction of
time the worker spends at the workstation
CO = 8-hr TWA breathing zone concentration at
workstation before recirculation
KB = fraction of worker’s breathing zone air that is
composed of recirculated air returning from
the air cleaner (range 0 to 1.0)
NOTE: Units for CB, CG and CO are in parts per million (ppm) or milligrams per cubic meter [mg/m3] and
all the values must be in the same unit system.
The coefficients KR, KB, and f are dependent on the workstation and the worker’s position in relation to the source of the
recirculated air returning from the air cleaner and in relation to
the exhaust hood. The value of KR can range from 0 to 1.0 where
0 indicates no recirculated air entering the hood and 1.0 indicates
100% of recirculated air entering the hood. Similarly, the value
of KB can range from 0 to 1.0 where 0 indicates there is no recirculated air in the breathing zone and 1.0 indicates that the breathing zone air is 100% recirculated air from the air cleaner. The
coefficient “f” varies from 0 where the worker does not spend
any time at the workstation where the air is being recirculated to
1.0 where the worker spends 100% time at the workstation. In
many cases, it will be difficult to attempt quantification of the
values required for solution of these equations for an operation
not yet in existence. Estimates based on various published and
other available data for the same or similar operations may be
useful. Furthermore, it may be difficult to determine accurate KR
and KB values for existing systems unless contaminant is easily
detected with direct-reading instruments or surrogate tracer gas
testing is used. Tracer gas testing will require highly trained
technicians. It is advised the final system be tested to demonstrate that it meets design specifications.
EXAMPLE PROBLEM 11-3 (Recirculation Calculations)
An example of use of Equations 11.2 and 11.3 and the effect of the
various parameters is as follows. Consider a system consisting of
5,000 acfm [2.5 am3/s] of general exhaust and 5,000 acfm [2.5
am3/s] of local exhaust. If the local exhaust is recirculated, the
ventilation system of QA = 10,000 acfm [5.0 am3/s] changes to
5,000 acfm [2.5 am3/s] recirculated plus QB = 5,000 acfm [2.5
am3/s] fresh airflow. Assume poor placement of the supply air
register(s) delivering the cleaned recirculated air back from the air
cleaner (KR and KB = 1) and that the worker spends all his time at
the workstation (f = 1); the air cleaner efficiency h = 0.90; exhaust
duct concentration (CE) = 500 ppm; general room concentration
(CG) = 20 ppm; replacement air concentration (CM) = 2 ppm;
workstation (breathing zone) concentration before recirculation
(CO) = 14 ppm; and a contaminant TLV® of 25 ppm.
Equation 11.2 gives recirculation air return concentration:
Initially:
QB = 10,000 [5.0], QA = 5,000 [2.5], CE = 500, CG = 20,
CM = 2, f = 1, Co = 14, KB = KR = 1
CR = [(1 – .9)(500 – (1)(2)] / [1 –( 1)(1 – .9)]
CR = [(.1)(498)] / [1 – .1] = 55.33 ppm = 55 ppm
Equation 11.3 gives the worker breathing zone concentration:
CB = (10,000/5,000)(20 – 2) (1 – 1) + (14 – 2)(1)
+ (1)(55) + (1 – 1)(2)
CB = (0) + 12 + 55 + 0 = 67 ppm
The concentration of 67 ppm is over the TLV® of 25 ppm and,
therefore, is not an acceptable value for an engineered
ventilation solution to this situation. Note that value is same in
IP and SI units.
In order to achieve lower concentrations (CB), two modifications in the design are made. First, the efficiency of the air cleaning device is improved to h = 97%. Second, the system configuration is redesigned so that only 30% of the recirculation return
air reaches the workstation. Thus, KR and KB are reduced to 0.3.
Substituting these new data in Equations 11.2 and 11.3, the concentration in the air stream leaving the air cleaner drops to 15.0
ppm and the breathing zone concentration calculates as 18 ppm.
This is less than the TLV® of 25 ppm and, therefore, the design
would normally be acceptable. While this performance is technically acceptable, the owner may choose to further redesign the
system to an even lower percentage of the TLV®. In many cases,
Supply Air Systems
an engineered design with the exposure below 50% of the TLV®
avoids administrative and medically related regulatory requirements.
After changing parameters: h = 0.97; KB = KR = 0.3
CR = [(1 – .97) (500 – .3(2))] / [1 – [(.3)(1 – .97)]]
CR = [(.03) (499.4)] / [1 – [(.3) (.03)]]
CR = [14.98] / [.991] = 15.12 ppm = 15 ppm
11-29
6) Routine testing, maintenance procedures, and records
should be developed for recirculating systems.
7) Provide periodic testing of the workroom air. Air cleaners should not be installed without proven (certifiable
performance) expectations, data-driven knowledge
regarding change-out and breakthrough and capability to
detect when breakthrough has occurred.
8) Design an appropriate sign in a prominent place, which
reads:
CB = (10,000/5,000) (20 – 2)(1 – 1) + (14 – 2)1
+ (.3)(15) + (1 – .3)(2)
CB = 0 + 12 + (4.5 + 1.4) = 17.9 ppm = 18 ppm
11.8.2 Design Considerations for Air Recirculation.
More requirements associated with the recirculation of
exhaust can be found in published standards. The American
National Standards Institute (ANSI) issued ANSI 9.7,
Recirculation of Air from Industrial Process Exhaust
Systems.(11.15)
Care must be taken in the design of recirculated air systems.
Considerations for good system performance are:
1) Recirculating systems should, whenever practicable,
be designed to bypass to the outdoors, rather than recirculate, when weather conditions permit. If a system is
intended to conserve heat in winter months and if adequate window and door openings permit sufficient
replacement air when open, the system can discharge to
the outdoors in warm weather. In other situations where
the workspace is conditioned or where mechanically
delivered supply air is required at all times, continuous
bypass operation may not be attractive.
2) Wet collectors also act as humidifiers. Recirculation of
humid air from such equipment can cause uncomfortably high humidity and require auxiliary ventilation or
some method to prevent excess humidity. Avoid the
return of overly humidified air into air conditioned
spaces.
3) The exit concentration of typical collectors can vary
with time. Consider design data and testing programs
that reflect all operational time periods.
4) The layout, design and delivery of the recirculated air
returning from the air cleaner should provide adequate
mixing with other supply air and avoid uncomfortable
drafts on workers or disruptive air currents which could
interfere with the capture velocity of local exhaust hoods.
5) Odors or nuisance values of contaminants should be
considered as well as recognized exposure limit values.
In some areas, adequately cleaned recirculated air, provided by a system with safeguards, may be of better
quality than the ambient outdoor air available for
replacement air supply.
CAUTION
AIR CONTAINING HAZARDOUS SUBSTANCES IS
BEING CLEANED TO A SAFE LEVEL IN THIS EQUIPMENT AND RETURNED TO THE BUILDING. SIGNALS OR ALARMS INDICATE MALFUNCTIONS
AND MUST RECEIVE IMMEDIATE ATTENTION:
STOP RECIRCULATION, DISCHARGE THE AIR OUTSIDE, OR STOP THE PROCESS IMMEDIATELY.
11.8.3 Recirculation Air Monitor Selection. While all system components are important, give special consideration to
the monitor on any system recirculating a potentially hazardous material. The prime requisites are that the monitor be
capable of sensing a system malfunction or failure, and of providing a signal that will initiate an appropriate sequence of
actions to assure that overexposure does not occur. The sophistication of the monitoring system can vary widely. The type of
monitor selected will depend on various parameters (i.e., location, nature of contaminant — including shape and size — and
degree of automation). The safe operation of a recirculating
system depends on the selection of the best monitor for a given
system. There are four basic components of a complete monitoring system, which include signal transfer, detector/transducer, signal conditioner, and information processor. Figure
11-19 shows a schematic diagram of the system incorporating
these four components. It is quite likely that commercially
available monitors may not contain all of the above four components and may have to be custom engineered to the need.
In addition to the four monitoring system components, the
contaminant samples must be collected from the air stream
either as an extracted sample or in total. If a sample is taken, it
must be representative of the average conditions of the air
stream at that point in time. At normal duct velocities, turbulence assures good mixing so gas and vapor samples should be
representative. For aerosols, however, the particle size discrimination produced by the probe may bias the estimated
concentration unless isokinetic conditions are achieved.
The choice of detection methods depends on the measurable
chemical and physical properties of the contaminants in the air
stream. Quantifying the collected contaminants is generally
much easier for particulate than for gases, vapors, or liquid
aerosols. If the exhaust air stream contains a toxic substance as
11-30
Industrial Ventilation
FIGURE 11-19. Schematic diagram of recirculation monitoring system
defined in the OSHA Hazard Communication Standard, an
exhaust air recirculation system must utilize a continuous
monitoring device. This device must be able to detect a concentration of the toxic substance of 10% of the acceptable
level of the contaminant as well as a reduction in exhaust airflow that is greater than 10% of the design.(11.15)
Particulates. Where the hazardous contaminant constitutes
a large fraction of the total dust, filter samples may allow adequate estimation of concentration in the recirculated air. If the
primary collector (e.g., bag filters, cartridge filters) allows
very low penetration rates, it may be more economical to use
high efficiency filters as secondary filters. If the primary filter
fails, the secondary filter not only will experience an easily
measured increase in pressure drop, but will filter the penetrating dust as well (Figure 11-