HANDBOOK OF
AIR CONDITIONING
SYSTEM DESIGN
OTHER McGR’AW-HILL
.‘\MERIC.\N
~NSTITI:TE
OF
P H Y S I C S
A~~IXIC.I\N
Socrli~v 01: MIXIIANICAL
Engineering Tables
Metals
Engineering-Design
Metals Engineering-Processes
Metals Properties
AMERICAN
SOCICTY
OF
TOOL
AND
HANDBOOKS OF INTEREST
. American Institute of Physics Handbook
ENGINLXRS . ASME Handbooks:
MANUFACTURING
ENGINEERS:
Die Design Handbook
Manufacturing Planning and
Estimating Handbook
Handbook of Fixture Design
Tool Engineers Handbook
ARCHITECTURAL RECORD
. Time-Saver Standards
BEEn<AN . Industrial Power Systems Handbook
BRADY . Materials Handbook
BURIKGTON . Handbook of Mathematical Tables and Formulas
BURINCTON AND MAY . Handbook of Probability and Statistics with Tables
CARROLL
Industrial Instrument Servicing Handbook
CONDON
AND ODISHAW . Handbook of Physics
CONSIDIN~
. Process Instruments and Controls Handbook
CONSIDINE
A N D R o s s Handbook of Applied Instrumentation
CROCKER Piping Handbook
D A V I S . Handbook of Applied Hydraulics
DUDLEY
Gear Handbook
EMERICK . Heating Handbook
FACTORY MUTUAL ENGINEERING DIVISION . Handbook of Industrial Loss Prevention
FL&XE . Handbook of Engineering Mechanics
HARRIS . Handbook of Noise Control
HARRIS AND CREDE . Shock and Vibration Handbook
HEYEL . The Foreman’s Handbook
HUSKEY AND KORN . Computer Handbook
JURAN . Quality Control Handbook
KALLEN . Handbook of Instrumentation and Controls
KING AND BRATER . Handbook of Hydraulics
K N O W L T O N . Standard Handbook for Electrical Engineers
KOELLE . Handbook of Astronautical Engineering
KORN AND KORN . Mathematical Handbook for Scientists and Engineers
LASSER . Business Management Handbook
LAUGHNER AND HARGAN . Handbook of Fastening and Joining of Metal Parts
LEGRAND
. The New American Machinists’ Handbook
MACHOL
System Engineering Handbook
MAGILL, HOLDEN, AND ACKLEY . Air Poilution Handbook
MAXAS . National Plumbing Code Handbook
MASTELL . Engineering Materials Handbook
MARKS AND BAUXEISTER
. Mechanical Engineers’ Handbook
MAYNARD . Industrial Engineering Handbook
MAYNARD
. Top Management Handbook
M O R R O W ’ Maintenance Engineering Handbook
PERRY . Chemical Engineers’ Handbook
PERRY . Engineering Manual
ROSSNACEL
Handbook of Rigging
ROTHBART
Mechanical Design and Systems Handbook
SHAND
Glass Engineering Handbook
STASIAR
Plant Engineering Handbook
STREETER Handbook of Fluid Dynamics
TOIJLOUKIAN
Retrieval Guide to Thermophysical Properties Research Literature
HANDBOOK OF
AIR CONDITIONING
SYSTEM DESIGN
Carrier Air Conditioning Company
.
&
I
e M
New York
C
GRAW-HILL
Sara Francisco
BOOK
Toronto
COMPANY
London
Sydney
HANDBOOK
apyright
@
OF
AIR
1965
by
CONDITIONING
McGraw-Hill,
Inc.
SYSTEM
All
Rights
DESIGN
Reserved.
@ 1960, 1963, 1964, 1965 by Carrier Corporation. Printed in the United States of
-4merica.
This book, or parts thereof, may not be reproduced in any form without
permission of the publishers. Lib7ary of Congress Catalog Card Number 65-17650.
ISBN
07-010090-X
15 16 17 18 HDHO 8543210
.
PREFACE
The Handbook of Air Conditioning System Design is the first complete
practical guide to the design of air conditioning systems. It embodies all the
knowledge and experience gained over the past fifty years by the pioneer in the
field, Carrier Air Conditioning Company.
This handbook is tailored to the specific needs of the man responsible for
the details of design, and, therefore, the foremost consideration has been the
requirements of the consulting engineer. In fact, many of the concepts embody
the up-to-date thinking of consulting engineers.
If any one word best describes this work, it is the word “practical.”
l
l
l
l
l
l
It is usable at all educational levels.
I
It provides practical data for professional designers who need optimum
solutions on a day-to-day basis.
It bridges the gap between air conditioning texts and manufacturers’
product catalogs.
It provides proved system design techniques and assures quality of application with minimum service requirements.
It provides guidance in simplified form.
It provides a reference source employing the best techniques of indexing
and format.
This Handbook of Air Conditioning System Design is a companion piece to
manufacturers’ product literature. Together the handbook and product literature make up a complete engineer’s manual.
Those using this book for study will benefit from clear applicable examples
presented in each of the engineering sections.
In summary, this Handbook of Air Conditioning System Design is a quick
reference for those actively engaged in designing’ air conditioning systems, a
teaching work for those studying air conditioning system design, and a refresher
for those engineers with wide experience in the field.
*
*
*
Grateful appreciation is hereby extended to those hundreds of Carrier engineers who generously contributed to the total body of knowledge herein, and
to those consulting engineers, mechanica contractors, and architects who so
willingly and enthusiastically contributed their experience to this project.
Carrier Air Conditioning Company
TENAGA
EWBANK PRt$CE,
LIBRARY
CONTENTS
p re f ace . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
P a r t 1 ..--..:LOAD--.--
ESTIMATING . .‘.
I-l
.:: . . . . . . . . . . . . .
Building Survey and Load Estimate . . . . . . . .
Design Conditions . . . . . . . . . . . . . . . . . . . . . . .
Heat Storage, Diversity and Stratification .
Solar Heat Gain thru Glass . . . . . . . . . . . . . .
Heat and Water Vapor Flow thru Structures
Infiltration and Ventilation . . . . . . . . . . . . . .
Internal and System Heat Gain . . . . . . . . . . .
Applied Psychrometrics
..................
1.
2.
3.
4.
5.
6.
7.
8.
V
.
.
.
. .
.
. .
. .
.
.
. . .
.
.
.
. . .
.
. .
l-l
l-9
l-25
l-41
l-59
l-89
1-99
l-115
1
,Part
lib_
.
2.~ AIR DISTRIBUTION :. . . . . . . . . . . . . . . . . . . .
1. Air Handling Apparatus . . . . . . . . . . . . . . . . . . . . . . . . .
2. Air Duct Design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
3. Room Air Distribution . . . . . . . . . . . . . . . . . . . . . . . . . . .
2-l
. . . . .
.
2-l
2-17
2-65
DESIGN .. :: . . . . . . . . . . . . . . . . . . . . . .
. . . .
......
3-1
Piping Design-General . . . . . . . . . . . . . . . . . . . . . . . . .
Water Piping . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
Refrigerant Piping . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
Steam Piping . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
. .
. . . .
......
......
.
3-l
3-19
3-43
3-81
. . .
: Part 4. REFRIGERANTS;
BRINES; OIL.‘?: : :. :. .
L.-.-.-..~
I
1. Refrigerants .
.
.
.
.
.
. . . .
. .
.
2. Brines . . . .‘... . . . . . . . . . . . . . . . . . . .
. . . .
3. Refrigeration Oils . . . . . . . .
.......
4-l
.......
.......
.......
4-l
4-23
4-55
--Part
:- -. 1.
2.
3.
4.
3 : PIPING
.+ Part 5. WATER CONDITIONING
..
1. %ter Conditioning-General
2. Scale and Deposit Control . . .
3. Corrosion Control . . . . . . .
4. Slime and Algae Control . . . .
5. Water Conditioning Systems .
6. Definitions . . . . . . . . . . . . . . . . .
. . . . . . . . . , . . . . . . . . .
5-1
. . . .
. . .
. . . .
5-l
5-l 1
5-19
5-27
5-31
5-47
‘._
.
.
. .
. .
.
.
. .
’
I
Part 6 . AIR HANDLING
EQUIPMENT
...........................
G-l ~
I. Fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
2. Air Coriclitioning Apparatus . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
3. Unitary Equiprncnt . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
4. :l(~rcssory Erlliiprncnt . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
6-I
G-17
G--l 5
G-5 1
Part 7. REFRIGERATION EQUIPMENT . . . . . . . . . . . . . . . . . . . . . . . .
7-l
1. Reciprocating Refrigeration Mnrhine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
2 . Centrifugal Rcfrigcrntion Machine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
. 3. Absorption Refrigeration Machine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
-1. CombinaGon r\hsorption-Centrifugal System . . . . . . . . . . . . . . . . . . . . . . . .
5. ITent Rejection Equipment . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
7-1
7-21
7-33
7-47
7-55
Part 8. AUXILIARY EQUIPMENT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
8-l
1.
2.
3.
4.
Centrifugal Pumps . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
Motors and Motor Controls . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
Boilers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
Miscellaneous Drives . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
8-l
8-17
8-5 1
8-61
Part 9. SYSTEMS AND APPLICATIONS . . . . . . . . . . . . . . . . . . . . . . . . .
9-l
1. Systems and Applications . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
9-l
Part 10. ALL-AIR SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
1.
2.
3.
4.
5.
6.
Convention& Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
Constant Volume Induction System 1 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
Multi-zone Unit System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
Dual-duct System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
Variable Volume; Constant Temperature System . . . . . . . . . . . . . . . . . . . . .
Dual Conduit System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
.
10-I
10-l
10-9
10-17
lo-25
10-35
10-41
\
Part 11.
AIR-WATER SYSTEMS ....................................
11-l
1. Induction Unit System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
2. Primary Air Fan-coil System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
11-l
1 l-23
Part 12. WATER AND DX SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
12-1
1. Fan-coil Unit System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
2. DX Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
12-1
12-11
Index
..............................................................
I-l
/
k
HANDBOOK OF
AIR CONDITIONING
SYSTEM DESIGN
l-l
Par-t 1
LOAD ESTIMATING
CHAPTER 1. BUILDING SURVEY AND LOAD ESTIMATE
The primary function of air conditioning is to
maintain conditions that are (1) conducive to
human comfort, or (2) required by a product, or
process within a space. To perform this function,
equipment of the proper capacity must be installed
and controlled throughout the year. The equipment
capacity is determined by the actual instantaneous
peak load requirements; type of control is deterI
xd by, the conditions to be maintained during
ped~ and partial load. Generally, it is impossible to
measure either the actual peak or the partial load
in any given space; these loads must be estimated.
It is for this purpose that the data contained in Part
1 has been compiled.
Before the load can be estimated, it is imperative that a comprehensive suruey De made to assure
accurate evaluation of the load components. If the
building facilities and the actual instantaneous load
within a given mass of the building are carefully
studied, an economical equipment selection and system design can result, and smooth, trouble free performance is then possible.
The heat gain or loss is the amount of heat instantaneously coming into or going out of the space.
The actual load is defined as that amount of heat
which is instantaneously added or removed by the
enllipment. The instantaneous heat gain and the
’‘s.
11 load on the equipment will rarely be equal,
because of the thermal inertia or storage effect of
the building structures surrounding a conditioned
space.
Chapters 2, 4, 5, 6, and 7 contain the data from
which the instantaneous heat gain or loss is estimated. Chapter 3 provides the data and procedure
for applying storage factors to the appropriate heat
gains to result in the actual load. Chapter 8 provides
the bridge between the load estimate and the equipment selection. It furnishes the procedure for establishing the criteria to fulfill the conditions required
by a given project.
The basis of the data and its use, with examples,
are included in each chapter with the tables and
charts; also an explanation of how each of the heat
gains and the loads manifest themselves.
BUILDING SURVEY
SPACE CHARACTERISTICS AND HEAT LOAD
SOURCES
An accurate survey of the load components of the
space to be air conditioned is a basic requirement
for a realistic estimate of cooling and heating loads.
The completeness and accuracy of this survey is the
very foundation of the estimate, and its importance
can not be overemphasized. Mechanical and archi-
tectural drawings, complete field sketches and, in
some cases, photographs of important aspects are
part of a good survey. The following physical aspects
must be considered:
1. Orientation of building - Location of the ’
space to be air conditioned with respect to:
a) Compass points -sun and wind effects.
b) Nearby permanent structures - shading
effects.
c) Reflective surfaces - water, sand, parking
lots, etc.
2. Use of space(s) - Office, hospital, department
store, specialty shop, machine shop, factory,
assembly plant, etc.
3. Physical dimensions of space(s) - Length,
width, and height.
4. Ceiling height - Floor to Hoor height, Hoor to
ceiling, clearance between suspended ceiling
and beams.
5. Columns and beams - Size, depth, also knee
braces.
6. Construction materials - Materials and thickness of walls, roof, ceiling, floors and partitions, and their relative position in the structure.
7. Surrounding conditions - Exterior color of
walls and roof, shaded by adjacent building
or sunlit. Attic spaces - unvented or vented,
gravity or forced ventilation. Surrounding
spaces conditioned or unconditioned - temperature of non-conditioned adjacent spaces,
such as furnace or boiler room, and kitchens.
Floor on ground, crawl space, basement.
8. Windows - Size and location, wood or metal
l-2
I’
9.
10.
11.
12.
13.
14.
sash, single or double hung. Type of glass single or multipane. Type of shadi-ng device.
Dimensions of reveals and overhangs.
Doors - Location, type, size, and frequency of
use.
Stairways, elevators, and escalators- Location,
temperature of space if open to unconditioned area. Horsepower of machinery, ventilated or not.
People - Number, duration of occupancy,
nature of activity, any special concentration.
At times, it is required to estimate the number
of people on the basis of square feet per person, or on average traffic.
Lighting - Wattage at peak. Type - incandescent, fluorescent, recessed, exposed. If the
lights are recessed, the type of air flow over
the lights, exhaust, return or supply, should
be anticipated. At times, it is required to estimate the wattage on a basis of watts per sq ft,
due to lack of exact information.
M o t o r s - Location, nameplate and brake
horsepower, and usage. The latter is of great
significance and should be carefully evaluated.
The power input to electric motors is not
necessarily equal to the rated horsepower divided by the motor efficiency. Frequently these
motors may be operating under a continuous
overload, or may be operating at less than
rated capacity. It is always advisable to measure the power input wherever possible. This
is especially important in estimates for industrial installations where the motor machine
load is normally a major portion of the cooling load.
APPl iances, business machines, electronic
equipment - Location, rated wattage, steam
or gas consumption, hooded or unhooded, exhaust air quantity installed or required, and
usage.
Greater accuracy may be obtained by measuring the power or gas input during times of
peak loading. The regular service meters may
often be used for this purpose, provided power
or gas consumption not contributing to the
room heat gain can be segregated.
Avoid pyramiding the heat gains from various
appliances and business machines. For example, a toaster or a waffle iron may not be used
during the evening, or the fry kettle may not
be used during morning, or not all business
!ART 1. LOAD ESTIMATING
machines in a given space may be used at the
same time.
Electronic equipment often requires individual air conditioning. The manufacturer’s
recommendation for temperature and humidity variation must be followed, and these requirements are often quite stringent.
15. Ventilation - Cfm per person, cfm per sq Et,
scheduled ventilation (agreement with purchaser), see Chapter 6. Excessive smoking or
odors, code requirements. Exhaust fans-type,
size, speed, cfm delivery.
16. Thermal storage - Includes system operating
schedule (12, 16 or 24 hours per day) specifically during peak outdoor conditions, permissible temperature swing in space during a
design day, rugs on floor, nature of surface
materials enclosing the space (see Chapter 3).
17. Continuous or intermittent operation Whether system be required to operate every
business day during cooling season, or only
occasionally, such as churches and ballrooms.
If intermittent operation, determine duration
of time available for precooling or pulldown.
LOCATION OF EQUIPMENT AND SERVICES ,
The building survey should also include information which enables the engineer to select equipment
location, and plan the air and water distribution
systems. The following is a guide to obtaining this
information:
1. Available spaces - Location of all stairwells,
elevator shafts, abandoned smokestacks, pipe
shafts, dumbwaiter shafts, etc., and spaces for
air handling apparatus, refrigeration machines, cooling towers, pumps, and services
(also see Item 5).
2. Possible obstructions - Locations 0E all electrical conduits, piping lines, and other obstructions or interferences that may be in the
way of the duct system.
3. Location of all jire walls and partitions Requiring fire dampers (also see Item Ih).
4. Location of outdoor air intakes - In reference
to street, other buildings, wind direction, dirt,
and short-circuiting of unwanted contaminants.
5. Power seruice - Location, capacity, current
limitations, voltage, phases and cycle, 3 or 4
wire; how additional power (if required) may
be brought in ,and where.
6. Water seroice - Location, size of lines, ca-
.
CHAP-I-El<
I .
ljacity, pressure, maximum temperature.
7. SLetlm .semice - Location, size, capacity, ternperature, pressure, type of return system.
8. I<efrigerntion, brine o r chilled waler (if f u r nished by customer)-Type of systen!, capacity,
temperature, gpm, pressure.
9. Arrhitect7lral characteristics of spnce - For
selection of outlets that will blend into the
s p a c e design.
10. Existing air conveying equipment wnd ducts For possible reuse.
11. Drains - Location and capacity, sewage disposal.
12. Control facilities - Compressed air source and
pressure,
electrical.
13. Foundation and support - Requirements and
facilities, strength of building.
So&d
and vibration control requirements _ I.
Kelation of refrigeration and air handling
apparatus location to critical areas.
15. Accessibility fo? moving equipment to the
final location - Elevators, stairways, doors,
accessibility from street.
16. Codes, local and national- Governing wiring,
drainage, water supply, venting’of refrigeration, construction of refrigeration and air
handling apparatus rooms, ductwork, fire
dampers, and ventilation of buildings in general and apparatus rooms in particular.
AIR CONDITIONING LOAD ESTIMATE
The air conditioning load is estimated to provide
the basis for selecting the conditioning equipment.
It must take into account the heat coming into the
sy-ace from outdoors on a design day, as well as the
.
being generated within the space, A design day
is defined as:
1. A day on which the dry- and wet-bulb temperatures are peaking simultaneously (Chapter 2,
“Design
1-3
l\LJII,DING SlJKVEY .\ND LO>\11 ESl‘IM.\I‘E
Conditions”).
2. A day when there is little or no haze in the air
to reduce the solar heat (Chapter 4, “Solar Heat
Gain Thru Glass”).
3. All of the internal loads are normal (Chapter
7, “Internal and System Heat Gain”).
The time of peak load can usually be established
by inspection, although, in some cases, estimates
must be made for several different times of the day.
Actually, the situation of having all of the loads
peaking at the same time will very rarely occur. To
be realistic, various diversity factors must be applied
10 sonic of the lmtl compo11ents; rcl’er to cl/crfiter 3,
“ljeat .~torage, ljiwrsity, and .\‘t1.nti/ir/ltion.”
The infiltratiotl ;1nt1 ventilation air quantities arc
estim;rtecl as dcscribetl in CilnpleY h.
I;ig. I illustrates 211 air conditioning lmtl cstimatc
form ;tntl is dcsigllctl to permit systematic load eVaI-
uation. This form contains the references identified
to the particular chapters of data and tables rcquirctl
to estimate the various load components.
OUTDOOR LOADS
The loads from outdoors consist of:
1. Tile slln rays entering windows - Table 15,
pages 44-49, and Table 16, pnge 52, provide
data from which the solar lieat gain through
glass is estimated.
The solar heat gain is usually reduced by
means of shading devices on the inside or outside of the windows; factors are contained in
Table 16. In addition to this reduction, all or
part of the window may be shaded by reveals,
overhangs, and by adjacent buildings. Chart I,
page 57, and Table 18, page 55, provide an ,
easy means of determining how much the
window is shaded at a given time.
A large portion of the solar heat gain is radiant
and will be partially stored as described in
Chapter 3. Tables 7 thru 11, pages 30-34, provide the storage factors to be applied to solar
heat gains in order to arrive at the actual
cooling load imposed on the air conditioning
equipment. These storage factors are applied to
peak solar heat gains obtained from Table 6,
page 29, with overall factors from Table 16,
page 52.
2. The sun rays striking the walls and roof
-
These, in conjunction with the high outdoor
air temperature, cause heat to flow into the
space. Tables 19 and 20, pages 62 and 63, provide equivalent temperature differences for
sunlit and shaded walls and roofs. Tables 21,
22, 23, 24, 25, 27, and 28, pages 66-72, provide
the transmission coefficients or rates of heat
flow for avariety of roof andwall constructions.
3. The air temperature outside the conditioned
space - A higher ambient temperature causes
heat to flow thru the windows, partitions, and
floors. Tables 25 and 26, pages 69 and 70, and
Tables 29 and 30, pages 73 and 74, provide
the transmission coefficients. The temperature
differences used to estimate the heat flow thru
these structures are contained in the notes after
each table.
l-4
l’.\Rl‘ I . LO,\11 I:S’I‘I,\I.\~I‘IN(~
HAP
I
TABLE REFERENCES
AREA OR
ITEM
G
3
&
4
G
L
A
S
~S
S
T
O
R
A
G
OF OPERATION
CONDITEE.-FEouroooR(OA, j Tau l - 3
1 6 . 1 7
!
2
SOFTX,
TE L
&
TRANS.
So
WALL
S”FTX
WALL
So
SQ
P 6 2
I
x:
X
TEL
SaFlX!
ROOF-SHADE”
20
4
, XI
II
XI
~63
PP
66.69
TBLS
/
27.28
OUTDOOR AIR
X i TBL 45 I -CIY,PLRSOH
p97
i
x p-Cr*iSa
FT
PEOPLE
90
FT
, Yl
I iTOLS 2 1 2 2 . 1
/ A23,24 “R 251
~ TBL1.9
FT
‘EMT,.
ATION
.5 ROOF
FT X 1
FT X 1
/
So F T X ’
JOOF-SUN
15
GAIN-WALLS
WALL
WALL
5
FOR
‘1:
id roi 15 CORR I
S o F T k’
-“-;
: --’PP 44-49 ,<I PP 4
4
.
4
9
S o FT X’
GLASS
--1 WITHOUT
GLASS
-1S T O R A G E
SKYLIGHT !
SOLAR
,““SS
2 TBL”
9.lO”Rll
1
FT X, wL9-3++ PP 5 2 . 5 4
SQ
E
TABLE REFERENCES
:STIMATE
FACTOR
/
SOLAR GAIN-GLASS
S O FT U T0~56&7 6
Wlill
L A S S
SUN G AIN OR
TEMP DIFF.
!
CluANTlrv
iAF
EF
=tEF
CFM
SWNCIHG
REYOLVlNc.
6
ALL
GAIN-EXCEPT
GLASS
So
PARTlTlON
FT X
Sa F
-CEILI N G
WALLS
D O O R S - PEOPLE
OPEN Doons-oooas
IWTIL.
‘RITI”I(
‘i
EXHAUST
TBLS 4 6 . 4 7 : P 98
FAN
FEET x TBL
CRACK
7 1 . 7 2
NOTE
X NOTFS.
T
So Fr X
FLOOR
Sa
INFILTRATiON
NOTE
I’
FT
ROOF
1
X
!
X
PP
- CFM
TBL 3 3
4 4
P 9 5
CFY/FT = -
NQTF
1
AIR
‘X
PP
73
EFFECTIVE
sEFw*;“E;T
ESHF
74
HEAT
G-42 p g2
-!%ZE-CFMo
3
TBL 6 5
_
ADP
E F F E C T I V__._...
E
Room
SENS.
EAT _
. ..~ HEFFECTIVE
ROOM T O T A L HE A T
C H A R T. FI G 3 3 P 1 1 6
=
P 145. O R P S Y C H
INDICATED ADP =
SELECTED ADP =
----F
DEHUMIDIFIED
INTERNAL
n
APPARATUS
APPARATUS DEWPOINT
1.08
X
THRU
P 76
p; TBL 29 :R 30
73.74
OUTDOOR
T;pLS629Sj&6
PP 6%70X
NO T E S
X
CFM
4
(i
= ~
n
VENTlLATlON
CFM ,NFlLTRATl”N
TRANS.
= ~
AIR
__
_-.-F
QUANTITY
P 121
(I--BF) X CT,,- F - Tm-F) =-F
TpBppL;;4&3
PEOPLE X
PEOPLE
HP
WATTS
POWER
oil K W
x 3,4
A P P LI A N C E S .
TBLS
E TC .
ADD,T,ONAL
HE
50-52 PP 1 0 1 - 1 0 3
GAINS TBLS 5 4 . 5 7
A T
So
STORAGE
X $;;,” kj;k;;
8
PP 107-109 X
(TEMP
’ %iM
3UTLETI
X TBL 5 3 P 10:
x TBLS 12.14.49
35 ,R la!-
_ LIGHTS
sue
SWING?
T EMP .
DIFF.
EFFECTIVE
1.08
I
F ACTOR
ROOM
_
-
CFMc
RISE
SENS.
HE A T
= -FIRM--(IUTLET
CFM n.
SUPPLY
FT X i TBL 1 4 P 38 1X (- T;$
RO
%iY
TOTAL
HE A T
TCYP
~$8,
I ”
1.08
SAFETY
SENS.
F
~-1.08 X
TOTAL
SUE
ROOM
x
O M
AIR
SENS H E
X
F
QUANTITY
A T
DESIRED
.
pz.7
CFM:
O,FF
%
P 113
R~JOM
SE.NSlBLE
H E A T
W
RESULTING ENT G LVG CONDITIONS AT APPARATUS
O
U T D O O R
AIR
NOTE 3
CFM
E F F E C T I V E
Y
NOTE
R O O M
LATENT
INFILTRATION
PEOPLE
PEOPLE
STEAM
P
121
BF
ETC .
ADDlTloNAL
HE A T
TBLS 5 0 . 5 2
GAINS
TRANS.
SUPPLY
PP
x
TBLS 1 4 . 4 8 :
X t,tOO
LEAKAGE
NOTE
3
LOSS
E N S I B L E
:
NOTES
NOTE 3
CORR B
X TBL 40 P Q
SUB T
LATENT
%
____~
GRAB
CFM x N
CFM XN”TE
R O O M
T,,,-TLDB-
C H A R T : TEWB- F. T LWS-F
1.
p;My~~-B”~~
(“B) T E M P E R A T U R E
Form
O T A L
2.
~;&MOISTURE
CO
3.
NORMALLY .
US E &CFM VENTILATION”
EYER. WHEN ,NF,LTRAT,“N
1s T o B E
DETERM,NE
.-CFM OUTDOOR AIR.”
AIR
O T E
EZO.
4.
W H E N I N F I L T R A T I O N IS N O T 1” BE O F F S E T . A N ” “ C F M
VENTILATION
IS L E S S T H A N “ C F M INFILTRATION.”
T H E N T H E E X C E S S INFILTRATI”
,s Acco”NT~oFOR HERE
LATENT
HEAT
H E A T
n
HEAT
1
F X cl--~121BF)
X
Z”a,~s
Without
HEAT
N T E N T
P 1 2 1 BF x a,68
x
T O T A L
I.08
X (1 --P 12lEF) Y 0 . 6 8
TEL 6 0
S U B TO T A L
FL, p ‘Q -+ KY:,,“, pdj,
G R A N D
Carrier Masthead
=
TBLS
E L O W
P 110
ROOM
RETURN
C H A R T 3 RETURN
DUFT
P 1 1 0 DUCT
P 112
Oh +
HEAT GA,”
% + LEAK. GAlN
With
CT,,-F-TT,,-F)
NOTES
X
ROOM
CFM X NOTE 2
OUTDOOR
LATENT:
X
(T em---F - Tm--F) =
PP 38.100
%
E F F E C T I V E
S
PSYCH.
GII/LI1 x 0 . 6 8
2
X NOTE ZGI,!-8
EFFECTIVE
8
T,,-F-t&+’
T,,-F+?BF X
FRO”
NOTE
101.103
P 113
DUCT
I R
W
DIFFERENCE
FROM
TOP
“F ESTIMAT
FROM
T
“F
TBL 58 P 1 0 9
S0.f~
F ACTOR
O”TD”“R A
EDB
1.08
LDB
X
--__.
SAFETY
X
H E A T
LB/HI x 1050
P 107
APPLIANCES.
VAPOR
X
HEAT
CFM
Noi~ 4
1F
S E N S I B L E
T O T A L
H E A T
Cwrier M”sthe”dForm
n
E5024.
FIG . 1 - AIR CONDITIONING LOAD ESTIMATE
.
(GR,LB)
DI
F F E R E N C E
O P
ESTIMAT
FOR “ C F M
OUTDOOR A I R . ” H o w
OFFSET. R E F E R T o PA G E 9 2 T
CHAPTER
I. I1UILDINC;
SlJKVliY
.\NI> LO,\D
4. The clir 71~ipr)~ ~MS~~W - i\ Iiighcr wpoi
pressure surrounding contlitionctl space ca~scs
water v a p o r t o liow tilt-u the l)uilding niaterials. This load is signilicant only in low dcw1)oint al~l~lications.
Tile data required t o estinlate this load is containccl in Table JO, p(/ge
Sf. In comlort applications, this load is neglected.
5 . The wield blorui,lg
trgtrilrst ~1 side of tlte I~rliltl-
irjg-1Vintl c;iuscs the outdoor air that is higlicr
in tempcraturc ant1 moisture content to infiltrate thru the cracks around the doors and
windows, resulting in locali~ctl scnsiblc ant1
latent heat gains. All or part of this infiltration
may be ofFset by air being introduced thru the
apparatus for ventilation purposes.
Cl/npler h
contains the estimating data.
Ozltdoor
niy usunlly
wquired
1-s
ESI‘IM;\TE
/(II. ueiltilntiotl
pz+oses - Outdoor air is usually necessary to
flush out the space and keep the odor level
down. This ventilation air imposes a cooling
and dehumidifying load on the apparatus bec a u s e t h e h e a t and/or moisture must be
removed. i\Iost air conditioning equipment
permits some outdoor air to bypass the cooling
surface (see Chapter 8). This bypassed outdoor
air becomes a load within the conditioned
space, similar to infiltration; instead of coming
thru a crack around the window, it enters the
room thru the supply air duct. The amount
of bypassed outdoor air depends on the type
of equipment used as outlined in Chnptey 8.
Table $5, page 97, provides the data from which
the ventilation requirements for most comfort
applications can be estimated.
’ “ie foregoing is that portion of the load on the
a; .anditioning equipment that originates outside
the space and is common to all applications.
INTERNAL LOADS
Chapter 7 contains the data required to estimate
the heat gain from most items that generate heat
within the conditioned space. The internal load, or
heat generated within the space, depends on the
character of the application. Proper diversity and
usage factor should be applied to all internal loads.
AS with the solar heat gain, some of the internal
gains consist of radiant heat which is partially stored
(as described in Chnpter 3), thus reducing the load
to be impressed on the air conditioning equipment.
Generally, internal heat gains consist of some or
all of the following items:
1. People - The human body thru metabolism
c
VI
generates heat within itself ant1 rcleascs
it by
radiation, convection, and evaporation I’rom
the su~l;~cc, mtl by coI1vection
arid evaporation
in the respiratory tract. The amount of heat
generated 2nd rcleasetl depends on surrountling tcniperaturc and on the activity level ol’ the
person, as listed in Ttrble fS, f-“1ge 100.
2. Liglct.9 - Illi~niinaiits convert electrical power
into light ant1 licat (refer to Cllclptel- 7). Some
of the heat is radiant and is partially stored
(see C/upter 3).
3 . Appliuuces
- Rcstaiirants, hospitals, laboratories, and some specialty shops (beauty shops)
have electrical, gas, or steam appliances which
release heat into the space. Tables 50 th?-u 52,
pnges 101-103, list the recommendecl heat gain
values for most appliances when not hooded. If
a positive exhaust hood is used with the appliances, the heat gain is reduced.
4. Electric calcfilnting machines - Kefer to manufacturer’s data to evaluate the heat gain from
electric calculating machines. Normally, not
all of the machines would be in use simultaneously, and, therefore, a usage or diversity
factor should be applied to the full load heat
gain. The machines may also be hooded, or
partially cooled internally, to reduce the load
on the air conditioning system.
5. Electric motors - Electric motors are a signifi-
cant load in industrial applications and should
be thoroughly analyzed with respect to operating time and capacity before estimating the
load (see Item 13 under “Space Characteristics and Heat Load Sozlrces”).
It is frequently
possible to actually measure this load in existing applications, and should be so done where
possible. Table 53, page 105, provides data for
estimating the heat gain from electric motors.
6. Hot pipes and tanks - Steam or hot water
pipes running thru the air conditioned space,
or hot water tanks in the space, add heat. In
many industrial applications, tanks are open
to the air, causing water to evaporate into the
space. Tables 54 thou 58, pages 107-109 provide data for estimating the heat gain from
these sources.
7. Miscellaneous sowces - There may be other
sources of heat and moisture gain within a
space, such as escaping steam (industrial cleaning devices, pressing machines, etc.), absorption
oE water by hygroscopic materials (paper, textiles, etc.); see Chapter 7.
In addition to the heat gains from the indoor
and outdoor sources, the air conditioning equipment and duct system gain or lose hcnt.. The fans
and pumps requirctl to distribute the air or water
thru the system add heat; heat is also added to
supply and return air ducts running thru warmer
or hot spaces; cold air may leak out of the supply
duct and hot air may leak into the return duct. The
procedure for estimating the heat gains from these
sources in percentage of room sensible load, room
latent load, and grand total heat load is contained
in Clmrt 3, ;b~ge I IO, and Tables 59 rind 60, pa.ges
111-113.
HEATING LOAD ESTIMATE
The heating load evaluation is the foundation for
selecting the heating equipment. Normally, the
ating load is estimated for the winter design
temperatures (Chnpter 2) usually occurring at night;
therefore, no credit is taken for the heat given off
by internal sources (people, lights, etc.). This estimate must take into account the heat loss thru the
building structure surrounding the spaces and the
heat required to offset the outdoor air which may
infiltrate and/or may be required for ventilation.
Chnpter 5 contains the transmission coefficients and
procedures for determining heat loss. Chapter 6 contains the data for estimating the infiltration air
quantities. Fig. 2 illustrates a heating estimate form
for calculating the heat loss in a building structure.
Another factor that may be considered in the
evaluation of the heating load is temperature swing.
Capacity requirements may be reduced when the
temperature within the space is allowed to drop a
few degrees during periods of design load. This, of
-ourse, applies to continuous operation only. Table
pnge 20, provides recommended inside design
conditions for various applications, and Table 13,
page 37, contains the data for estimating the possible capacity reduction when operating in this
manner.
The practice of drastically lowering the temperature to 50 F db or 55 F db when the building is
unoccupied precludes the selection of equipment
based on such capacity reduction. Although this type
of operation may be effective in realizing fuel economy, additional equipment capacity is required for
pickup. In fact, it may be desirable to provide the
additional capacity, even if continuous operation is
contcmplatctl, because of pickup required after
forcctl shutdown. It is, thcrcl’ore, evident that the
use of storage in reducing the heating load for the
purpose of equipment selection should be applied
with care.
HIGH ALTITUDE LOAD CALCULATIONS
Since air conditioning load calculations are based
on pounds of air necessary to handle a load, a
decrease in density means an increase in cfm required to satisfy the given sensible load. The weiglit
of air required to meet the latent load is decreased
because of the higher latent load capacity of the
air at higher altitudes (greater gr per lb per degree
tlitference in dewpoint temperature). For the same
dry-bulb and percent relative humidity, the wctbulb temperature decreases (except at saturation)
as the elevation above sea level increases.
The following adjustments are required for high
altitude load calculations (see Chapter 8, Table 66,
page 148):
Design room air moisture content must be
adjusted to the required elevation.
.
Standard load estimating methods and forms
are used for load calculations, except that the
factors affecting the calculations of volume
and sensible and latent heat of air must be
multiplied by the relative density at the particular elevation.
Because of the increased moisture content of
the air, the effective sensible heat factor must
be corrected.
EQUIPMENT SELECTION
After the load is evaluated, the equipment must
be selected with capacity sufficient to offset this load.
The air supplied to the space must be of the proper
conditions to satisfy both the sensible and latent
loads estimated. Chapter S, “Applied PsychrometTics ,” provides procedures and examples for determining the criteria From which the air conditioning
equipment is selected (air quantity, apparatus dewpoint, etc.).
.
(:II.\I”TTR I. I:1711.1>1NC;
SllRVI<Y ,\ND I>O.\D
HEATING CONDITIONS
l-7
ESTIM,\'I‘I~
TEMPERATURE OF AIR ENTERING UNIT
TOTAL TRANSMISSION LOSS
FORM El0
F1c.2 - HEATING LOAD ESTIMATE
l-9
CHAPTER 2. DESIGN CONDITIONS
This chapter presents the data from which the
outdoor design conditions are established for various
localities and inside design conditions for various
applications. The design conditions established determine the heat content of the air, both outdoor
and inside. They directly affect the load on the
air conditioning equipment by influencing the
transmission of heat across the exterior structure
and the difference in heat content between the outdoor and inside air. For further details, refer to
OUTDOOR DESIGN CONDITIONS - SUMMER
AND WINTER
The outdoor design conditions listed in Table 1
are the industry accepted design conditions as published in AR1 Std. 530-56 and the 1958 ASHAE
:. The conditions, as listed, permit a choice of
G
ouLdoor dry-bulb and wet-bulb temperatures for different types of applications as outlined below.
B.NORMAL DESIGN CONDITIONS - SUMMER
Normal design conditions are recommended for
use with comfort and industrial cooling applications
where it is occasionally permissible to exceed the
design room conditions. These outdoor design conditions are the simultaneously occurring dry-bulb
and wet-bulb temperatures and moisture content,
which can be expected to be exceeded a few times
a year for short periods. The dry-bulb is exceeded
more frequently than the wet-bulb temperature, and
usually when the wet-bulb is lower than design.
When cooling and dehumidification (dehydration) are performed separately with these types of
applications, use the normal design dry-bulb tem-
perature for selecting the sensible cooling apparatus; use a moisture content corresponding to the
normal design wet-bulb temperature and 80% rh
for selecting the dehumidifier (dehydrator).
Daily range is the average difference between the
high and low dry-bulb temperatures for a 24-hr
period on a design day. This range varies with local
climate conditions.
A, MAXIMUM DESIGN CONDITIONS-SUMMER
Maximum summer design conditions are recommended for laboratories and industrial applications
where exceeding the room design conditions for
even short periods of time can be detrimental to a
product or process.
The maximum design dry-bulb and wet-bulb
temperatures are simultaneous peaks (not individual
peaks). The moisture content is an individual peak,
and is listed only for use in the selection of separate cooling and dehumidifying systems for closely
controlled spaces. Each of these conditions can be
expected to be exceeded no more than 3 hours in
a normal summer.
NORMAL DESIGN CONDITIONS - WINTER
Normal winter design conditions are recommended
for use with all comfort and industrial heating applications. The outdoor dry-bulb temperature can
be expected to go below the listed temperatures a
few times a year, normally during the early morning hours. The annual degree days listed are the
sum of all the days in the year on which the daily
mean temperature falls below 65 F db, times the
number of degrees between 65 F db and the daily
mean temperature.
’
l-10
PART I. LO/\D
ES?‘IM,\?‘ING
TABLE l-OUTDOOR DESIGN CONDITIONS-SUMMER AND WINTER
STATE
AND
CIT Y
ALABAMA
Anniston
Birmingham
Mobile
Montgomery
ARIZONA’
MAXIMUM
DESIGN
COND.-SUMMER
July at 3:00 PM
DI r y I
Bulb
(F)
DryBulb
(F)
95
95
95
95
90
105
65
76
81
94
26
30
TlJCSOll
105
100
110
72
70
78
-
77
85
93
30
95
95
76
78
-
104.5
117.5
I6
I6
25
ARKANSAS
Fort Smith
little Rock
CALIFORNIA
Bakersfield
El Centro
Eureka
Fresll0
Laguno Beach
Long Beach
Los Angeles
Oakland
Montague
Pasadena
Red Bluff
Sacramento
San Bernadine
San Diego
Son
Francisco
San Jose
Williams
105
110
90
105
70
78
65
74
54
94
52
76
90
90
85
70
70
65
78
78
60
9 5
100
100
70
70
72
70
62
73
I8
I05
85
85
72
68
65
65
75
60
IO
17
91
70
76.5
Durango
95
95
Fort Collins
Grand Junction
Pueblo
95
95
CONNECTICUT
Bridgeport
Hartford
New Haven
Waterbury
64
65
65
65
-
75
95.9
82
70
103.0
94
94
68
88
74
102
68
86.2
I4
16
14
117.5
Apalachicola
Jacksonville
Key West
Miami
95
95
98
91
80
78
78
79
I31
117.5
I 12.5
I31
Pensacola
Tampa
Tallahorsee
95
95
78
117.5
117.5
80
-
78.4
24
25
78
7.7 SW
5.4 E
6,894
1,108
5.0 w
5.2 NW
2,376
4,853
I46
1036
6.7 N
7.0 E
6.0 NW
30
25
4758
2403
7.0 N
8.0 NW
35
30
1391
6.0 SW
74.4
30
2680
35
35
1596
3137
25
2823
-
94
95
0
0
0
- 1 5
82
7.0 w
2.0 w
mtiuda
deg)
8.3 E
8.3 NW
448
324
7.3
5.4 NW
499
43
132
287
6.4 NE
84
155.6
99
82
150.5
92
81
150.5
34
34
31
32
32
35
33
10
47
261
I7
34
34
34
38
2.635,
42
34
40
39
7.2 SE
305
I16
6.3 NW
7.5 N
26
I7
34
33
38
100
86
37
39
7.0 s
7.5 s
5.22 I
6,558
40
37
5613
5558
6.0 SE
4.4 NW
7.9 NW
4,587
4,770
41
39
38
6113
5880
7.0 s
7.0 s
8.7 NW
9.4 N
9
58
23
41
42
41
42
N W
134
40
7.8 NW
72
39
IO.0 S W
99
733
694
IO
293
5.0 w
0
COLUMBIA
Washington
7242
1441
3226
3009
99.3
62
63
95
OF
110
89.4
117.5
-10
25
8.0 N
9.9 N
7.5 NW
25
68
78
1
83
145.5
-
99
95
2806
261 I
1566
207 I
25
-10
30
2.5
99
102
99
DIST.
126.9
60
70
75
75
75
Wilmington
78
155.6
110
95
93
95
DELAWARE
2
ilavathan
4bove
Se0
Level
(ft)
I
40
-
COLORADO
Denver
r
103
103
DATA
Avg. Velocity and
DryAnnual
Prevailing
Direction
~(pr/lb o f I B u l b I D e g r e e I
Summer
Winter
I c i,; a i r ) 1 IFI 1 Doyr
8
30
WIND
ontentt
Bulb i
VI
90
II3
1
lolrture
-Dry-
19
I9
12
I5
Flagstaff
Phoenix
Winslow
Yuma
‘_
Vetiulb
IF)
-
I
NORMAL
DESIGN COND.
WINTER
0
5.0 s
FLORIDA
78
95
‘Correspondr to dry-bulb and wet-l
3 temperatures listed, and is corrected for altitude of city.
tcorrespondr to peek dewpoint temperature, corrected for altitude.
25
25
45
35
20
30
25
-
1252
II85
59
I85
1281
571
1463
-
5 . 0 SW,
8.0 SW
9.0 SE
7.0 SE
8.4
9.0 NE
10.6 NE
10.1 E
23
I8
23
II
30
30
25
26
6.0 NE
10.9 N
8.6 NE
N
408
25
31
68
28
30
(:I
I.\I”I’I~:I<
11. I>I<,SI(;N
1-11
(:ONI)I’I‘IONS
TABLE 1 -OUTDOOR DESIGN CONDITIONS-SUMMER AND WINTER (Contd)
NORMAL DESIGN
COND.-SUMMER
July at 3:00 PM
STATE
AND
CITY
MAXIMUM DESIGN
COND.-SUMMER
July at 3:00 PM
NORMAL
DESIGN COND.
WINTER
WIND
Elevation
Above
Sea
Level
Winter i
vtt)
DATA
GEORGIA
Atlanta
Augusta
Brunswick
Il.7 NW
6.5 NW
975
195
NW
9.5 NW
408
42
Columbus
Macon
Savannah
6.7
IDAHO
Boise
Lewiston
Pocotello
fwin Falls
‘-
‘NOIS
3iro
Lhicogo
Danville
Moline
Peoria
Springfield
INDIANA
Evansville
Fort Wayne
Indianapolis
95
95
95
1
I
ii
54.5
44
65
65
65
/ ;i
9698
96
31
28
28
61
/ ‘it” 1 I9
76 76 77
71
92.6
100
1 I04
/ 80
1 140.6
-10
5678
5
- 5
-IO
5109
674 1
--I--
5.0 N W
9.1 SE
4.1 E
8.9 SE
/ -I: / :;:: /
10.0
763
4,468
9.8
12.0 SW
NW /
NE
8.3 s
I I.9 NW
I9
20
18
102
100
99
82
150.5
0
-IO
- 1 0
4410
7.0 S W
6232
8.0 S W
5458
9.0 S W
- 5
117.5
I23
117.5
I8
I8
- 5
-I5
-I5
- 2 0
102
6252
6375
6820
78
78
I25
132
20
21
‘opeko
,Vichita
100
100
78
75
109.5
98
I9
21
95
78
1 17.5
22
99
I
I I
LOUISIANA
Alexandria
New Orleans
Shreveport
I
5425
Ill
-IO
-IO
- 1 5
110
-IO
-IO
5075
106'
79
i
I
126.9
38
41
40
42
40
42
42
43
I
I
00
I
637
1,111
I
4417 4792
I 7.0 S W
1201
i
39
38
13.3 SW
9.8 SW
I
42
41
43
43
39
38
39
5069
4644
46
43
42
41
41
40
6.0 S W
10.0 s
11.0 s
44
594
602
603
8.2 SW
11.5 N W
95
95
33
33
32
37
42
40
+
Fort Dodge
Keokuk
Sioux City
Waterloo
KANSAS
Concordia
Dodge City
Salino
34
34
31
319
594
124
IOWA
Cedar Rapids
Davenport
Des Moines
Dubuque
KENTUCKY
Lexington
Louisville
2,705
W
I03
106 103
117.5
99
104.5
South Bend
Terre Haute
109
Latitude
(deeI
N
8.6 N
8.8 SE
989
459
38
38
89
9
197
32
30
33
45
45
44
Belfast
Eortport
Millinocket
Presque Isle
Portland
Rumford
I
90 I
70 I
781131
I
/
I 5: I
- 2 0
I
*Corresponds to dry-bulb and wet-bulb temperatures listed, and is corrected for altitude of city.
tcorresponds to peak dewpoint temperature, corrected for altitude.
8445
I 7.0 s
12.6 W
/
NW
10.4 NW
100
44
45
46
L
47
1
47
44
44
’
l-12
I’AKT
I .
LO:\11 IiS’I‘lhl,\l‘lNG
TABLE l-OUTDOOR DESIGN CONDITIONS-SUMMER AND WINTER (CONT.)
l-
STATE
AND
CITY
MAXIMUM
DESIGN
COND.-SUMMER
July at 3:00 PM
DryBulb
(F)
MARYLAND
Baltimore
Cambridge
Cumberland
_
_
.-Frederick
Frostburg
Salirbury
WetBulb
(F)
95
78
95
75
DryBulb
(F)
117.5
DryBulb
(F)
18
MICHIGAN
Alpeno
Big Rapids
Detroit
Escano ba
75
75
95
75
95
75
-I
9>
9.5
95
Ludington
Marquette
Saginaw
Soult Ste Marie
1
‘3
-10
0
-10
96
-10
-15
0
z+-
93
93
i
75
95
1 75
17
17
19
101
20
99
20
+
L
j ‘11
93
73
95
75
~__
St. Cloud
St. Paul
/
I
79
4487
90
99
96
103
20
20
-10
-10
96
I
Summer / Winter
6.0 SW
1 8.2 NW
NW
I
I
19
17
-2.5
-25
-20
102
t
79
131.1
+
MISSOURI
Columbia
Kansas City
Kirksville
St. Louis
St. Joseph
Springfield
79
78
100
100
78
76
95
/
70
18
98
20
104
82
L__-
/
66
‘55.6
-I
71
56
Gc
49
20
97
1
77.4
c
I
temperature, corrected for altitude.
199
625
42
42
42
42
615
45
I
8278
6560
8777
t
WI
6702
c
111.0 SW
NW
10.0 SW 12.0 SW
/9.5 NW
7149
t
‘1.9 w
, 10.6 NW
/
8.9 SE
1,128
7975
9.0 s
7213
7.7 SE
6.3 N
8.3
8.9 SW
10.3 NW
SW 1
9.0s
8.0 S
111.8s
9.3 NW
‘0.9 SE
j 12.4 W
84’6
t
i
7930
8032
7591
7604
47
*
5070
4962
-25
-20
-20
-30
;a
w
9723
7966
4596
5596
4569
j
, 9.8 SW
9307
c
6’9
8.0 w /12.’ Lzw
7458
8745
-10
-IO
*Corresponds to dry-bulb and wet-bulb temperatures listed, and is corrected for altitude of city.
peak dewpoint
9.0 SW
5.0 SW
4.0 SW
6.0 SW
0
0
5
0
40
40
40
W
2330
2069
2
2
3
2
39
39
39
-__.-~
6743
15
10
10
-
14
.
-
70
Latitude
kg)
c
0
-IO
-10
108
66
95
/
109
‘08
78
--
I
96
c
90
83
-25
-20
Above
Sea
Level
(ft)
5936
c
-20
103
95
95
tion
Avg. Velocity and
Prevailing
Direction
W
NW
-10
-I5
-IO
-I5
-10
-10
- 5
-IO
98
t
MlSSlSSlPPl
Jackson
teridian
ucksburg
135.9
c
/
93
73
95
75
MINNESOTA
Alexandria
Duluth
Minneapolis
102
102
0
- 5
-IO
0
99
75
75
I
EleVa-
DATA
L
0
5
-/‘02
17
_-
Flint
Grand Rapids
Kalamazoo
Lansing
tCorresponds to
AntWa
Degree
Days
WIND
L-.--.
104
.ew Bedford
/lymouth
Springfield
Worcester
Helena
Kalispell
Miles City
Misroula
B u l b (gr/lb o f
(F)
dry
air)
r
- 5
- 5
10
MASSACHUSETTS
Amherst
Boston
Fall River
-__
Fitchburg
Lowell
Nantucket
MONTANA
Billings
Butte
Great Falls
Hovre
Moisture
Net- contentt
NORMAL
XSIGN C O N D .
WINTER
5.6 S
E
3’6
410
226
32
32
32
( : I l.\l”1’l~:I~
1-13
2. I)lL‘5l(;N C:ONI)I~I‘IONS
TABLE 1 -OUTDOOR DESIGN CONDITIONS-SUMMER AND WINTER (CONT.)
NORMAL DESIGN
COND.-SUMMER
July at 3:00 PM
STATE
AVG.
DAILY
RANGE
MAXIMUM DESIGN
CON&-SUMMER
July at 3:00 PM
NORMAL
DESIGN COND.
WINTER
WIND DATA
ElevaLatitud
kg)
NEBRASKA
Grand Island
41
41
42
41
41
43
North Platte
Omaha
Valentine
York
NEVADA
tar Vegas
RetlO
Tonopah
‘%memucca
h
HAMPSHIRE
Berlin
Concord
115
95
75
65
40
41
95
65
40
90
73 i
102
66
66.9
20
- 5
5
- 1 5
562 I
5812
6357
7.0 S W
7.0 S W
s
6.0 W
9.9 SE
8.1 NE
1,882
4,493
5.42 1
4,293
I
I
I
95
j
14
I
I I
I /
I 17’
36
40
30
42
45
43
43
43
43
I
5
Camden
East Oronae
+
Newark Cit;
Jersey
Paterson
Sandy Hook
Trenton
95
95
95
75
75
95
99
99
78
a
14
14
117.5
99
95
14
82
145.5
0
81
140.6
0
0
5015
13.0 S W
15.8 N W
10.0 S W
5500
96
13.0 S W
13.0 S W
9.0 S W
NW
17.1 N W
16.1
10.9 N W
125
30
173
10
10
56
39
41
40
41
41
41
41
41
40
35
32
36
77
_ lens F a l l s
Ithaca
Jamestown
Lake Placid
Roche&r
Schenectady
Syracuse
I
I
I
I
126.9
- 5
- 2 5
- 1 0
- 1 5
1
+
I
1
I
95
93
93
/
I
I
75
75
75
I
102
I
18
I
95
6925
8305
12.0 S W 1 17.1 w
8.0
10.5
NW
W
604
458
43
42
43
43
43
43
42
42
- 44 41
45
i 43
43
43
43
43
44
36
35
37
36
34
‘Corresponds to dry-bulb and wet-bulb temperatures listed, and is corrected for altitude of city.
tcorresponds to peak dewpoint temperature, corrected for altitude.
i
TABLE 1 -OUTDOOR DESIGN CONDITIONS-SUMMER AND WINTER (CONT.)
NORMAL DESIGN
COND.-SUMMER
MAXIMUM DESIGN
COND.-SUMMER
AVG.
DAILY
NORMAL
DESIGN COND.
STATE
AND
CITY
NORTH DAKOTA
Birmarck
Devils Lake
c
Fargo
Grand Forks
Williston
OHIO
Akron
Cincinnati
Cleveland
Columbus
Dayton
Lima
Sandusky
Toledo
Youngstown
OKLAHOMA
Ardmore
Bartlesville
Oklahoma City
Tulsa
95
95
95
75
78
75
99
117.5
99
95
95
76
78
104.5
123
19
22
19
23
23
95
95
95
75
75
75
99
99
99
19
19
106
101
81
79
- 5
0
0
145.5
135.9
95
99
99
4990
6144
7.0 SW
11.0 s
8.5 SW
14.7 SW
104
553
651
41
39
42
-10
0
- 5
5506
5412
9.0 SW
8.0 SW
11.6 SW
11.1 SW
724900
40
40
41
0
-IO
6095
6269
10.0 SW
II.0
12.1 SW
608
589
1,186
I
101
101
77
77
108
101.5
21
I
104
106
42
42
41
i
79
t
OREGON
Baker
Eugene
Medford
Pendleton
Portland
Roseburg
Wamic
1
/
I
-5
-15
/
7197
I
(
I
1
I
366
/
44
PENNSYLVANIA
Altoono
Bethlehem
Erie
Harrisburg
New Castle
Oil City
Philadelphia
Pittsburgh
Reading
Scranton
Warren
~ Williamsport
95
95
95
75
78
75
99
117.5
105
18
14
14
97
98
95
95
75
75
99
99
14
95
RHODE ISLAND
Block Island
Pawtucket
Providence
95
93
93
75
75
75
99
102
102
SOUTH CAROLINA
Charleston
Columbia
Greenville
95
95
95
78
75
76
117.5
99
104.5
17
17
17
98
SOUTH DAKOTA
Huron
Rapid City
Sioux Falls
95
95
95
75
70
75
106
05
99
19
22
20
I06
103
/
*Corresponds to dry-bulb ond wet-bulb temperatures listed,
tcorrespondr
to peak dewpoint
79
126.9
0
0
0
- 5
-15
- 5
4739
5430
10.0 SW
9.OSW
5232
6218
6.0 SW
15
IO
10
1066
2488
3059
42
40
40
11.0 NW
11.6 W
26
1,248
9.0
7.6 SW
NW
NW
311
746
525
40
41
41
42
10.5 SW
8.0 SW
8.4
9
401
902
33
34
35
14
14
a2
155.6
76
71
and is corrected for altitude of city.
temperature, corrected for altitude.
10.0 SW
7.0 N E
,
,- _ ----_-.., _.
l-15
(:f I \l”l~I~:l< 2. I)LSI~;N (:ONl)I-I‘IoNS
TABLE l-OUTDOOR DESIGN CONDITIONS-SUMMER AND WINTER (CONT.)
STATE
AND
CITY
TENNESSEE
Chattanooga
Johnson CiLy
-Knoxville
Memphis
Nashville
TEXAS
Abilene
Amarillo
Austin
Brownsville
Corpus Chrirti
Dallas
Del Rio
El Po’so
Fort Worth
Galveston
Houston
Palestine
Port Arthur
San Antonio
UTAH
Modem
Logan
Ogden
Salt take City
VERMONT
Bennington
Burlington
Rutland
DryBulb
(F)
Net.
Bulb
WI
95
76
95
95
95
75
78
78
100
100
100
-.- 95
95
100
100
100
100
95
95
100
95
100
l-I-onirteunrat*
MAXIMUM
DESIGN
COND.-SUMMER
July at 3100 PM
gr/lb of
dry air)
DryBulb
VI
DryBulb
(F)
104.5
I8
98
I7
103.5
100
117.5
18
103
117.5
17
98
---i--
74
72
78
80
80
78
78
69
78
80
80
78
79
78
124
109.5
95..
65
95
65
90
90
73
73
WetBulb
(F)
NORMAL
DESIGN COND.
WINTER
Vloisture
htentt
:gr/lb of
dry air)
Avg. Velocity and
Annual tI
Degree
Days
Summer
I
6.0 S W
79
83
7.7
NW
ElSVotlon
Above
Sea
Level
(ft)
/
689
6.0 S W
7.0 S W
8.0 w
Lotitude
(deeI
35
- 36 36
35
36
-
93
101
96
I
105
80
101
72
100
81
I9
102
83
66
25
97
61
25
102
2573
4196
1679
628
965
2367
1501
2532
2355
II74
1315
2068
1532
1435
9.0 s
11.0 s
9.0
13.0
8.0
10.0
9.0
10.0
9.0
8.0
32
35
31
26
28
33
29
32
33
29
30
32
30
21
SE
SE
s
SE
E
s
s
S
7.8 SE
68
-
91
8.0 s
11.6 S
4,446
4,222
38
42
41
41
308
VIRGINIA
Cape Henry
Lynchburg
Norfolk
Richmond
Roancke
vVASHINGTON
North Head
Seattle
Spokane
Tacoma
Tatoosh Island
Walla Wallo
Wenatchee
Yakima
WEST VIRGINIA
Bluefield
Charleston
Elkinr
Huntington
Martinsburg
Porkersburg
Wheeling
95
95
95
95
95
78
75
78
78
76
99
95
90
85
85
93
85
65
65
65
64
86
106
95
90
95
65
65
65
37
37
37
38
38
11.0 s
6.0 SW
5367
70
68
7.0 N
7.0 SW
I05
W
102
76
98
*Corresponds to dry-bulb and wet-bulb temperatures listed, and is corrected for altitude of city.
tcorrerponds to peak dewpoint temperature, corrected for altitude.
16.1
9.8 SE
6.2 S W
8.0
18.9
5.4 s
4928
4.0 SE
48
48
47
48
46
48
47
603
37
38
39
38
39
39
40
,
TABLE l-OUTDOOR DESIGN CONDITIONS-SUMMER AND WINTER (CONT.)
STATE
AND
CITY
NORMAL DESIGN
COND.-SUMMER
July at 3:00 PM
Moisture
Wet- content*
Bulb (gr/lb of
(F) dry air)
DryBulb
(F)
WISCONSIN
Ashland
Eau Claire
Green Bay
L o Crosre
Madison
Milwaukee
WYOMING
Carper
Cheyenne
Lander
Sheridan
AVG.
DAILY
RANGE
95
95
95
95
75
75
75
75
99
99
103.5
99
95
95
65
65
68.5
66
DryBulb
(F)
MAXIMUM
DESIGN
COND.-SUMMER
July at 3:00 PM
Moisture
vi.,- contentt
Bulb (gr/lb of
(F) dry air)
DryBulb
IF)
14
/
99
-r 17 100
18
96 ,
14
99
28
28
79
131.1
161.2
83
102
NORMAL
DESIGN COND.
WINTER
DryBulb
(F)
WIND
DATA
Avg. Velocity and
Annual Prevailing Direction
Degree
Days
Summer
Winter
-20
- 2 0
- 2 0
- 2 5
-I5
- 1 5
7931
742 I
7405
7079
8.0
6.0
8.0
9.0
-20
-15
-18
- 3 0
7536
8243
7239
9.0 s
5.0 S W
5.0 N W
S
S
SW
SW
10.5
9.3
10.1
12.1
Elevation
Above
Sea
Level
(ft)
SW
NW
SW
s
I
NW
w
885
589
673
938
619
SW
13.3 N W
3.9
4.9 N W
5,321
6,139
5,448
3,773
PROVINCE
AND
CITY
Lethbridge
MCMUrKly
Medicine Hat
BRITISH COLUMBIA
Estevon Point
Fort Nelson
Penticton
Prince George
Prince Rupert
VClllCOtJVC3~
I
NEW BRUNSWICK
Campbellton
Fredericton
Moncton
Saint John
NEWFOUNDLAND
Corner Brook
Gander
Goose Bay
Saint Johns
43
42
44
45
90
90
-
I
66
71
68
77
-
90
65
80
67
90
71
90
75 I
9520
9
3
10320
9
2 ’ 8650
2
5
8650
2
3
3
3
4
3
i
78
-
83.5
107
/
I
9.7
8.9
10.1
7.6
7.9
15.0
9.1
9.0
3
4
3
2
2
2
9
9
9500
6910
5230
5410
10930
16810
10630
-11
- 6
8830
-8 -3
8700 8380
-I
- 3
- 2 6
1
9210
9440
12140
8780
NORTHWEST
TERRITORIES
Aklovik
Fort Norman
Frobirher
Resolute
Yellowknife
*Corresponds to dry-bulb and wet-bulb temperatures listed, and is corrected for altitude of city.
tcorrerponds to peak dewpoint temperature, corrected for altitude.
3,540
2,219
2,190
3,018
1,216
2,365
51
54
55
50
57
50
7.2
8.0
7.7
12.3
20
1,230
1,121
2,218
170
22
228
49
59
50
54
54
49
48
14.7
6.4
12.0
1,200
115
894
786
50
59
54
50
9.2
42
164
48
46
13.8 14.9
248 119
46 45
17.2
10.3
19.3
40
482
144
463 ,
49
49
53
48
30
300
68
56
682
69
65
9.9
3.7
l7
- 3 8
- 6
- 3 2
8
11
15
Victoria
MANITOBA
Brandon
Churchill
The Pas
Winnipeg
42
45
45
44
43
43
.
CANADA
ALBERTA
Calgary
Edmonton
Grand Prairie
Latitude
be)
11.5
7.9
9.2
.
62
1-17
(:fl.\l”l’l-I< II. I)ICSI(iN (:ONI)I’I’IONS
TABLE 1 -OUTDOOR DESIGN CONDITIONS-SUMMER AND WINTER (CONT.)
NORMAL DESIGN
COND.-SUMMER
July at 3:00 PM
CANADA
PROVINCE
AND
CITY
COND.-SUMMER
DryBulb
VI
Direction
f------I
----I
NORMAL
DESIGN COND.
WINTER
DryBulb
VI
WIND
DATA
Avg. Velocity and
‘reVallill
Annual
Degree
Day5
Summer
Winter
Elevation
Above
SW
Level
(ft)
Latitude
idee)
NOVA SCOTIA
90
Halifax
Sydney
Yarmouth
ONTARIO
Fort William
Hamilton
Kapuskodng
Kingston
Kitchener
&don
North Boy
OttOWCl
Peterborough
Souix Lookout
Sudbury
Timminr
Toronto
Windsor
Soult Ste. Marie
PRINCE
ISLAND
-
75
107
7570
8220
7520
Ia
+--I
90
75
107
93
75
102
.93
75
102
- 2 4
0
- 3 0
-11
- 3
-I
- 2 0
-IS
-11
- 3 3
- 1 7
- 2 6
0
3
,’
,.-’
9.6
13.1
13.5
83
197
136
45
46
44
T
8.4
9.6
10.0
644
303
752
11.9
11.3
912
1,210
48
43
49
44
43
43
46
45
44
50
47
- 7810 7380
8830
9.6
8.9
7020
8.1
8380
a.7
=---I+
,48
43
42
47
+
EDWARD
Charlottetown
,:’
/’
10350
6890
11790
9.9
6.6
- 3
H QUEBEC
Arvido
Knob Lake
Mont Joli
Montreal
Port Harrison
Quebec City
Seven Islands
Sherbrooke
Three Rivers
90
90
75
107
90
75
107
-~__
1
- 1 9
- 4 0
-II
- 9
- 3 9
- 1 2
- 2 0
- 1 2
- 1 3
10440
-I
8130
9.9
9070
9.0
74
8.2
375
46
55
48
46
58
47
50
45
46
8610
t
SASKATCHEWAN
Prince Albert
Regina
Soskatoon
Swift Current
11.3
90
90
90
90
71
70
70
92.5
81
-41
- 3 4
- 3 7
- 3 3
11430
10770
10960
9660
YUKON
TERRITORY
DaV4?.0n
Whitehorse
*Corresponds to dry-bulb and wet-bulb temperatures listed, and is corrected for altitude of city.
tcorrespondr to peak dewpaint temperature, corrected for altitude.
12.4
10.7
4.9
12.1
9.7
14.6
15040
1
8.7
1,414
1,884
1,645
2,677
53
50
52
50
1,062
2,289
64
61
.
1-18
I’.‘IlR-I‘
CORRECTIONS TO OUTDOOR DESIGN CONDITIONS
FOR TIME OF DAY AND TIME OF YEAR
The normal design conditions for summer, listed
in Table 1, are applicable to the month of July at
about 3:00 P.M. Frequently, the design conditions
at other times of the day and other months of the
year must be known.
I. LOAD ESTIhI.\?‘ING
Solution:
Normal design conditions for New York in
July at 3:00
p.m. are 95 F db, 75 F wb (Table I).
Daily range in New York City is 14 F db.
Yearly range in New York City = 95 - 0 = 95 F db.
Correction for time of day (12 noon) from Table 2: I
Dry-bulb = -5 F
Wet-bulb = -1 F
Correction for time of year (October) from Table 3:
Table 2 lists the approximate corrections on the
dry-bulb and wet-bulb temperatures from 8 a.m. to
12 p.m. based on the average daily raqge. The drybulb corrections are based on analysis of weather
data, and the wet-bulb corrections assume a relat i v e l y c o n s t a n t dewpoint t h r o u g h o u t t h e 24-hr
period.
Dry4,ulb
= -16 F
Wet-bulb = - 8 F
Design conditions at 12 noon in October (approximate) :
Dry-bulb = 95 - 5 - 16 = 74 F
Wet-bulb = 75 - 1 - 8 = 6 6 F
INSIDE COMFORT DESIGN CONDITIONS SUMMER
Ta6le 3, lists the approximate corrections of the
dry-bulb and wet-bulb temperatures from March to
November, based on the yearly range in dry-bulb
I
3erature (summer normal design dry-bulb minus
w,,lter nprmal design dry-bulb temperature). These
corrections are based on analysis of weather data
and are applicable only to the cooling load estimate.
The inside design conditions listed in Table 4 are
recommended for types of applications listed. These
conditions are based on experience gathered from
many applications, substantiated by ASHAE tests.
The optimum or deluxe conditions are chosen
where costs are not of prime importance and for
comfort applications in localities having summer
outdoor design dry-bulb temperatures of 90 F or less.
Since all of the loads (sun, lights, people, outdoor
air, etc.) do not peak simultaneously for any prolonged periods, it may be uneconomical to de’sign
for the optimum conditions.
Example J - Corrections to Design Conditions
Given:
A comfort application in New York City.
Find:
The
approximate dry-bull, and wet-bulb temperatures at
12:00 noon in October.
TABLE 2-CORRECTIONS IN OUTDOOR DESIGN TEMPERATURES FOR TIME OF DAY
(For Cooling Load Estimptes)
DAILY
YGE OF
’ .-MPERATURE*
(F)
10
15
my
i../
25
30
35
40
45
/
. _ /,;UN
.,,
DRYOR
WETBULB
B
Dry-Bulb
W&-Bulb
Dry-Bulb
Wet&lb
- 9
- 2
-12
- 3
Dry-Bulb
Wet-Bulb
Dry-Bulbs
Wet-Bulb
Dry-Bulb
Wet-Bulb
-14
- 4
-I6
- 4
-18
- 5
-10
- 3
-IO
- 3
-12
- 3
- 5
- 1
- 5
-1
- 6
- 1
Dry-Bulb
Wet-Bulb
Dry-Bulb
Wet-Bulb
Dry-Bulb
Wet-Bulb.
-21
- 6
-24
- 7
-26
- 7
-14
- 4
-16
- 4
-17
- 5
- 7
- 2
-8
- 2
-
AM
PM
10
/
-
12
7
2
9
2
TIME
-.
I
-5- 1
-5,
->1 \
-a
- 2
,....+:-I
0
- 1
0
- 1
0
- 1
0
- 1
0
- 1
0
- 1
0
- 2
0
I
3 -A
0
0
0
0’
00
0
0
0
9
d
1)
0
0
0
0
4
I
I
-1
0
- 1
0
-1 0
6
I
- 2
- 1
- 2
-1
-3-1
-5
- 1
-6
- 1
-7 -2
-1
0
-I
0
-
3
1
4
1
-8
- 2
-1Q
- 3
- a
- 2
-10
- 3
-11
- 3
-13
- 3
-15
- 4
- 1
0
-1
0
- 2
-I
- 6
- 1
- 7
- 2
-8
- 2
-12
- 3
-14
- 4
-16
- 4
-18
- 5
-21
- 6
-24
- 0
* T h e d a i l y range of dry-bulb temperature is the difference between the highest and lowest dry-bulb temperature during o
d e s i g n day. (See Table I for the value of daily range for a particular city).
Equation: Outdoor design temperature at ony time = Outdoor design temperature from
10
8
I
I
24-hour
Table I + Correction from above table.
-
12
I
- 9
- 2
-14
- 4
-16
- 4
-18
- 5
-21
- 6
-24
- 7
-20
- 9
-31
-10
period on a typical
(:fI.\I”I’b:I~
2.
l-19
I)l3I(;N (:oKl)I~l‘I<)NS
TABLE 3-CORRECTIONS
IN OUTDOOR DESIGN CONDITIONS FOR TIME OF YEAR
Cooling
(For,
YEARLY
RANGE OF
TEMPERATURE(Fj*
120
TIME
March
I
90
/
OF
YEAR
April
Drv-Bulb
- 3 0
---IS
95
Estimates)
DRY- OR
WET-BULB
_ii
100
Load
;;j
~
‘-;i
zj+;
__~~.-I_
- 4
20
-I1
10
-5
- 2
_I-
Wet-Bulb
- 3 0
-15
Dry-Bulb
Wet-Bulb
Dry-Bulb
Wet-Bulb
- 2 9
- - 1 4
- 2 9
-14
-19
-10
-19-p
-IO
-10
-5
-
Dry-Bulb
Wet-Bulb
-29
-14
-19
-10
-IO
-5
- 2 9
-14
-19
-10
-9
-5
a5
Dry-Bulb
Wet-Bulb
80
Dry-Bulb
Wet-Bulb
75
Dry-Bulb
Wet-Bulb
70
Dry-Bulb
Wet-Bulb
65
Dry-Bulb
Wet-Bulb
60
Dry-Bulb
Wet-Bulb
55
Dry-Bulb
Wet-Bulb
50
Dry-Bulb
Wet-Bulb
--
-20
-IO
-11
-5
- 4
- 2
-IO
-5
3
2
3
2
0
0
0
0
e~wIqge~
-I_. - 0
I
-6
0
-3
0
-b
0
-3
$
;i
-I7
- B
-331--16
-17
- 2 9
-14
2;
- a
- 2 7
-14
- 2 7
-I4
- a
0
0
0
0
1;
0
0
0
0
-6
-3
-6
-3
- 1 6
-3
- 2
0
0
0
0
-16
- B
-26
-14
- 3
- 2
0
0
0
0
-5
-3
- 1 6
-25
- 1 4
j
- a
,,
‘Yearly range of temperature is the difference between the wmmer and winter normal desigq dry-bulb temperatures (Table 1).
Equation: Outdoor design temperature = Outdoor design temperature from Table 1 + Correct& from above table.
The commercial inside design conditions are recommended for general comfort air conditioning ap
plications. Since a majority of people are comforta!2’
t 75 F or 76 F db and around 45y0 to 50% rh,
tht --Ltermostat is set to these temperatures, and these
conditions are maintained under partial loads. As
the peak loading occurs (outdoor peak dry-bulb and
wet-bulb temperatures, 100% sun, all people and
lights, etc.), the temperature in the space rises to thi
design point, usually 78 F db.
If the temperature in the conditioned space is
forced to rise, heat will be stored in the building
mass. Refer to Chapter 3, “Heat Storage, Diversity
and Stmtification,” for a more complete discussion
of heat storage. With summer cooling, the temperature swing used in the calculation of storage is the
difference between the design temperature and the
normal thermostat setting.
The range of summer inside design conditions is
provided to allow for the most economical selection
of equipment. Applications of inherently high sen-
sible heat factor (relatively small latent load) usually
result in the most economical equipment selection if
the higher dry-bulb temperatures and lower relative
humidities are used. Applications with low sensible
heat factors (high latent load) usually result in more
economical equipnient selection if the lower drybulb temperatures and higher relative humidities
are used.
INSIDE COMFORT DESIGN CONDITIONSWINTER
For winter season operation, the inside design
conditions listed in Table 4 are recommended for
general heating applications. With heating, the
temperature swing (variation) is below the comfort
condition at the time of peak heating load (no
people, lights, or solar gain, and with the minimum
outdoor temperature). Heat stored in the building
structure during partial load (day) operation reduces
the required equipment capacity for peak load operation in the same manner as it does with cooling.
f
’
i
PAR-l- I
l-20
.
LO,\D E!xIhI,\‘-rIKG
TABLE 44tECOMMENDEDINSIDE DESIGN CONDITIONS*-SUMMER AND WINTER
WINTER
SUMMER
With
Commercial Prodice
TYPE OF
APPLICATION
Humidification
Without
Humidification
GENERAL COMFORT
Apt., House, Hotel, Office
Hospital, School, etc.
RETAIL
SHOPS
(Short term occupancy)
Bank, Barber or Beauty
Shop, Dept. Store,
Supermarket, etc.
LOW SENSIBLE HEAT
FACTOR APPLICATIONS
(High Latent load)
Auditorium, Church, Bar,
Restaurant, Kitchen, etc.
FACTORY COMFORT
Assembly Areas,
Machinina Rooms. etc.
I
I
I
I
I
,
* he room design dry-bulb temperature should be reduced when hot radiant panels ore adjacent to the occupant and increased when cold Panels are
~di,,en+ to compensate for the increase or decrease in radiant heat exchange from the body. A hot IX Cold Panel m’JY be unshaded glass or 9l“ss
block wjidows (hot in summer, cold in winter) and thin partitions with hot or cold spaces adjacent. An unheated slab floor on the ground or walls below
the ground level are cold panels during the winter and frequently during the summf~ ~Iso. Hot tanks, fUrnoceS Or machines are hot Panels.
TTemperoture swing ir above the thermostat setting at peak summer load conditionsJTemperature
swing is below the thermostat setting at peak winter load conditions (no lights, People Or solar heat gain).
**Winter humidification in retail clothing shops is recommended to maintain the quality texture of goods-
INSIDE INDUSTRIAL DESIGN CONDITIONS
Table 5 lists typical temperatures and relative
humidities used in preparing, processing, and manufacturing various products, and for storing both raw
and finished goods. These conditions are only typical
of what has been used, and may vary with applications. They may also vary as changes occur in
processes, products, and knowledge of the effect o f
temperature and humidity. In all cases, the temperature and humidity conditions and the permissible limits of variations on these conditions should
be established by common agreement with the customer.
Some of the conditions listed have no effect on the
product or process other than to increase the efficiency oE the employee by maintaining comfort
conditions. This normally improves workmanship
and uniformity, thus reducing rejects and production cost. In some cases, it may be advisable to
compromise between the required conditions and
comfort conditions to maintain high quality commensurate with low production cost.
Generally, specific inside design conditions are
required in industrial applications for one or more
of the following reasons:
1. A constant temperature level is required for
close tolerance measuring, gaging, machining, or grinding operations, to prevent expansion and contraction of the machine parts,
machined products and measuring devices.
Normally, a constant temperature is more im-;
portant than the temperature level. A constant:
relative humidity is secondary in nature but’,
should not go over 457, to minimize formation
of heavier surface moisture film.
Non-hygroscopic materials such as metals, glass,
plastics, etc., have a property of capturing
water molecules within the microscopic surface
crevices, forming an invisible, non-continuous
surface film. The density of this film increases
when relative humidity increases. Hence, this
film must, in many instances, be held below a
critical point at which metals may etch, or the
electric resistance of insulating materials is significantly decreased.
2. Where highly polished surfaces are manufac-tured or stored, a constant relative humidity
and temperature is maintained, to minimize
increase in surface moisture film. The temperature and humidity should be at, or a little
(:F1.\l”l’lCK
2.
I)ESI(;N (:ONI)I’I’IONS
below, the comfort conditions to minimize
perspiration of the operator. Constant temperature ant1 humidity may also be r~quirctl
in machine rooms to prevent etching or corrosion of the parts of the machines. With
applications of this type, if the conditions are
not maintained 24 hours a day, the starting of
air conditioning after any prolonged shutdown
shoultl bc tlonc carefully: (1) During the summer, the moisture accumulation in the space
should be reduced before the temperature is
reduced; (2) During the winter, the moisture
should not be introduced before the materials
have a chance to warm up if they are cooled
during shutdown pcriotls.
3. Control of relative humidity is required to
maintain the strength, pliability, and regain of
hydroscopic materials, sucll as textiles and
pAper. The humidity must also be controlled
in some applications to reduce the effect of
static electricity. Development of static electric
charges is minimized at relative humidities of
55% or higher.
4. The temperature and relative humidity control are required to regulate the rate of chemical or biochemical reactions, such as drying of
l-21
varnishes or sugar coatings, preparation of
synthetic fibers or chemical compounds, fermentation of yeast, etc. Generally, high temperatures with low humidities increase drying
rates; high temperatures increase the rate of
chemical reaction, and high temperatures and
relative humidities increase such processes as
yeast fermentations.
5. Laboratories require precise control of both
temperature and relative humidity or either.
Roth testing and quality control laboratories
are frequently designed to maintain the ASTM
Standard Conditions’ of 73.4 F db and 50%
rh.
6. With some industrial applications where the
load is excessive and the machines or materials
do not benefit from controlled conditions, it
may be advisable to apply spot cooling for the
relief of the workers. Generally, the conditions
to be maintained by this means will be above
normal comfort.
*Published in ASTM pamphlet dated 9-29-48. These conditions have also been approved by the Technical Committee
on Standard Temperature and Relative Humidity Conditions’
of the FSB (Federal
Specifications Board) with one variation: FSB permits 24%/,, whereas ASTM requires *2710 permissable humidity tolerance.
l-24
I’,\RT I. LOAD ESTIMATING
TABLE S-TYPICAL INSIDE DESIGN dONDlTlONS-INDUSTRIAL
(Listed
conditions
INDUSTRY
are
only
PROCESS
ABRASIVE
rtanufacture
BAKERY
)ough M i x e r
:ermen~ing
‘roof Box
bread Cooler
Zold Room
&cake-up Rm.
Zake Mixing
lrackerr & Biscuih
Nrapping
jtorogeDried
typical;
r
Ingred.
design
75-80
45-50
75-80
75-82
92-96
70-80
40.45
78-82
95-105
60-65
60-65
40-50
70-75
80-85
80-85
65-70
50
60-65
70
55-65
45-70
55-60
80
35
Water
32-35
-
Wax Paper
70-80
40-50
Sugar
are
established
50-65
CERAMICS
30-32
Grain
80 .
32-34
Lager
32-35
75
Ale
40-45
75
40-45
75
55
75
32-35
75
80-85
60-65
75-80
40-50
50-55
55-60
Ale
Racking
CANDYCHOCOLATE
Cellar
Candy Centers
Hand Dipping Rm.
Enrobing Rm.
EnrobingLoading End
Enrober
Stringing
Tunnel
Packing
Pan Specialty Rm.
General
Storage
CANDY-HARD
CHEWING GUM
Mfg.
Mixing 8 Cooling
TlIllIlel
Packing
storage
D r y i n g - J e l l i e s , Gum
C o l d Rm.Marshmallow
Mfg.
Rolling
Stripping
Breaking
Wrapping
80
90
70
40-45
65
70-75
65-/O
75-80
75-80
55
65-75
65-75
120-150
-
-
50
13
,
40-50
1DP - 40
30-40
40-45
DP - 55
40-45
45-50
15
75-80
45-50
77
68
72
74
74
33
63
53
47
58
PROCESS
efroctory
\olding Rm.
lay storage
lecol 8 Decorating
DRYBULB
(Fl
110-150
80
60-80
___
75-80
ackoging
75-80
Ifa.
65-70
DISTILLING
w a g e Grain
ELECTRICAL
PRODUCTS
lectronic & X - r a y
Coils & Trans.
Winding
ube Arrem.
lectrical Inst.
M f g . 8. t a b .
‘hermortot Assem. &
Calib.
humidistat Assem.
6. &lib.
3o.w Tol. Assem.
Aeter Asrem. Test
iwi~chgear-
-_
60
--__
Liquid Yeast
32-34
M-f g .~~~~~___ 60-75
~- 45-60
Aging
65-72
50-60
Fuse 8. Cut-Out
Assem.
CCID. W i n d i n a
Paper storage
Conductor
Wrapping
.ightning Arrestor
Circuit Brkr.
Assem. & Test
?ectifiersProcess Selenium
& Copper Oxid<
Plates
Fermenting CellarLager
requirements)
COSMETICS
55-60
~60
___75
Liauid Yeasl
customer
CEREAL
storageHops
by
INDUSTRY
80-85
Shortening
conditions
DRYBULB
(F)
30-45
-.
70-75
Fresh Ingred.
BREWERY
final
FURS
Drying
Shock Treatmeni
Storage
GLASS
Zutting
Vinyl Lam. Rm.
LEATHER
D r y i n g -
72
68
70
76
76
72
74-76
50-55
40-45
60-63
73
50
73
-7 3
75
68
50
50
65-70
20-40
76
30-60
74
30-40
110
18-20
40-50
55-65
Con
‘leg. Tanned
Chrome Tanned
storage
LENSESOPTICAL
Fusing
Grinding
MATCHES
Mfg.
D rying
storage
MUNITIONS
MetalPercussion
Elementr-
80
72-74
70-75
Drying Parts
D&a Points
Black Powder Drying
Condition & Load
Powder Type Fuse
toad Tracer Pellets
,
Comfort
I
50
”
50
40
(;1-1,\1”1‘1:11
2.
l)I,.sI(;s
(:oIUI)l’I’IoNS
1-23
TABLE 5-TYPICAL INSIDE CONDITIONS-INDUSTRIAL (Contd)
(Listed
conditions
INDUSTRY
PHARMACEU.
TICAL
PHOTO
MAT E R I A L
only
Powder
design
DRYBULB
(F)
.
70-80
30-35
conditions
are
TEXTILES
15-35
-75-80~~.
80
3570-80
1 40
80
35
-
PRECISION
MACHINING
Spectrographic Anal.
sear Matching 8.
Assem.
jtorageGasket
Cement 8 Glue
Aachinings
Gogin& A s s e m .
Adjusting Precision
Ports
Honing
REFRIGERATION
EQUIPMENT
Valve Mfg.
Compressor Assem.
Refrigerator Asrem.
resting
RUBBER DIPPED
GOODS
Mfg.
Cementing
Surgical Articles
storage Before Mfg.
.ob. (ASTM std.)
:otton
O p e n i n g 8. P i c k i n g
Cording
D r a w i n g 6. R o v i n g
requirements)
DRYBULB
(F)
PROCESS
Cotton,
cont.
Ring Spinning
Conventional
-
-__ -
80-85
80-85
80-85
60-70
78-80
78-80
75
60-65
70-85
65-70
75
55-65
55-60
C___-.___
arding, Spinning
75.80
60
W
eaving
80
80
__--.-.
-_. ___...WOOlWlS
~-Pickers
80-85
60
~.._._ _ __----Carding
80-85
65-70
~__- -_ __.Spinning
80-85
50-60
~-.._
--~__
_____
Dressino
75-80
60
-.
Weaving~___
__tight Goods
80-85
55-70
~__-_
- - - -.-.
Heavy
80-85
60-65
-__~---_
Drawing
75
50-60
-__.._ ___
I
W o~-rrteds
- _-..
C a r d i n g , C o m b i n g I,
B Giliing
80-85
60-70
___storage
70-85
75-80
~~--..
Drawing
80-85
50-70
___Cap Spinning 80-85
50-55
__- ~S p o o l i n g , Winding I
75-80
--__ - 55-60
Weavino
80
50-60
A---Finishing
75-80
60
-__
silk
.-. ~. _
Prep. 8. Dressing -.
80
60-65
Weavinq
8.
Spin&a
80
65-70
Throwina
80
60
__-.- _
?ayon
-__
Spinning
80-90
50-60
-__
Throwing
80
55-60
.__--_
Weovina
--___Regenerated
80
50-60
--__
Acetate
8
0
55-60
--.-_
Spun Rayon
80
80
~Picking
75-80
50-60
Carding, Roving,
Drawing
80-90
50-60
__Knitting
- -~
--.
Viscose or
Cuprammonium
80-85
65
__~.
Synthetic Fiber
Prep. & Weaving
Viscose
80
60
___--_
_ _ _
CdO”eS.2
80
70
25-30
~__
45-65
60
15-25
Comfort
_1
35-45
M u l t i c o l o r Litho.
Pressroom
Stockroom
Sheet & W e b P r i n t .
Storage, Foldino, etc.
customer
Cloth
__Room
-.-Combing
Linen
70
30-50
80
40
78-80
5-10
80
35
78
40-50
75
35-40
Comfort
80
35
70-80
20-30
Comfort
Comfort
90
90
(cont.)
by
Long Draft
Frame Spinning
S p o o l i n g , W o r p i n 19
_Weaving
Film _)
Hfg.Thermo S e t t i n g
Compounds
Cellophane
established
INDUSTRY
Drying
C u t t i n g & Packing
storoge~_~-_
Film Base, Film
Paper, Coated
Paper
Safety Film
lot Press-Resin
:old Press
TEXTILES
final
Storage
Before Mfg.
After Mfa.
Milling Rm.
Tablet Compressing
Tablet Coating
Effervescent~-__Tablet 8. Powder
Hypodermic Tablet
Colloids
Cough Syrup
Glandular Prod.
Ampule Mfg.
Gelatin Capsule
Capsule
storage
Microanalysis
Biological Mfg.
Liver Extract
Serums
Animal Rm.
PLYWOOD
PRINTING
typical;
PROCESS
rlitrote
PLASTIC
are
Comfort
Comfort
Nylon
TOBACCO
C i g a r 84 Cigarette
Mfg.
Softening
S t e m m i n g 8 Strippin 9
storage 8. Prep.
Conditioning
Qcking
8. S h i p p i n g
80
50-60
70-75
90
75-85
78
75
75
55-65
85.88
75
70
75
60
l-25
CHAPTER 3. HEAT STORAGE, DIVERSITY
AND S’JXATIFTCAT~ON
7%~ i~or111;~1 load cstini;tling proccdurc has been
to evaluate the instanlancous heat gain to a space
and to assume tllat the equilmlcnt will I;cmovc the
hcnt at this rate. Generally, it was found tllat the
ec~~~ipwmt sclccted
on this basis was
oversized and
therefore cnpablc of maintaining much lower room
conditions than the original design. Extensive annlysis, research and testing have shown that the reasons
for this arc:
1. Storage of heat in the building structure.
9 Non-simultaneous occurrcncc of the peak of
the individual loads (diversity).
3. Stratification of heat, in some cases.
This chapter contains the data and procedures
for determining the load the equipment is actually
picking LIP at any one time (actual cooling load),
taking into account the above factors. Application
of these data to the appropriate individual heat
gains results in the actual cdbling load.
The actual cooling load 1s generally considerably below the peak total instantaneous heat gain,
thus requiring smaller equipment to perform a
specific job. In addition, the air quantities and/or
water quantities are reduced, resulting in a smaller
overall system. Also, as brought out in the tables,
if the equipment is operated somewhat longer during the peak load periods, and/or the temperature
in the space is allowed to rise a few degrees at
the peak periods during cooling operation, a further
action in required capacity results. The smaller
system operating for longer periods at times of peak
load will produce a lower first cost to the customer
with commensurate lower demand charges and
lower operating costs. It is a well-known fact that
equipment sized to more nearly meet the requirements results in a more efficient, better operating
system. Also, if a smaller system is selected, and is
based on extended periods of operation at the peak
load, it results in a more economical and efficient
system at a partially loaded condition.
Since, in most cases, the equipment installed to
perform a specific function is smaller, there is less
margin for error. This requires morexexacting engineering including air distribution design and
system balancing.
With multi-story, multi-room application, it is
usu;~lly tlcsirable to provide some flcsibility in the
air sitlc or room load to ;~llow I’or individual room
control, load pickup, etc. Ccncrnlly, it is recommended that the full retl~lction from storage and
diversity be taken on the overall rcfrigcrntion or
I~iiiltling load, with some degree of conservatism on
the air side or room loads. This degree should be
determined by the cnginccr from project requiremcnts and customer desires. A system so designed,
full reduction on refrigeration load and less than
full reduction on air side or room load, meets all
of the flcsibility requirements, except at time of
peak load. In addition, such a system has a low
owning and operating cost.
STORAGE OF HEAT IN BUILDING
STRUCTURES
The instantaneous heat gain in a typical comfort
application consists of sun, lights, people, transmission thru walls, roof and glass, infiltration and
ventilation air and, in some cases, machinery, appliances, electric calculating machines, etc. A large
portion of this instantaneous heat gain is radiant
heat which does not’ become an instantaneous
load on the equipment, because it must strike a solid
surface and be absorbed by this surface before becoming a load on the equipment. The breakdown
on the various instantaneous heat gains into radiant
heat and convected heat is approximately as follows:
HEAT GAIN SOURCE
Solar, without inside blinds
Solar, with inside blinds
Fluorescent
Lights
Incandescent
Lights
People*
Transmission+
Infiltration
and
Ventilation
Machinery or AppIianceQ
RADIAN? * C O N V E C T I V E
HEAT
H E A T
100%
58%
50%
80%
40%,
6’3%
-
20.80%
42%
50%
20%
20%
40%
100YO
80s200(,
*The remaining 40% is dissipated as latent load.
+Transmission
load is considered to he 100yO convective load.
This load is normally a relatively small part of the total load,
and for simplicity is considered to I)e the instantaneous
load
on the equipment.
$Thc
load from machinery or appliances varies, depending
upon the temperature of the surface. The higher the surface
temperature, the greater the radiant heat load.
CONSTANT SPACE TEMPERATURE AND EQUIPMENT
OPERATING
PERIODS
,,\s tlic r;ttli:\nt
heat from S O L I I ‘ C ~ S s h o w n i n t h e
above table strikes ; I solid surface
(walls, floor, cciling, etc..), it is absorbed, r a i s i n g the tcmperaturc
at
the surface of the n~aterinl above that insitlc the
material and the air adjacent to the surt’ace. This
temperature ditfcrcncc causes heat flow into the
material by conduction and into the air by convection. The heat conducted away from the surface is
stored, and the heat convected lrom the surface
becomes an instantaneous cooling load. The portion
of radiant heat being stored depends on the ratio
of the resistance to heat flow into the material and
the resistance to heat flow into the air film. With
most construction materials, the resistance to heat
flow into the material is much lower than the air
resistance; therefore, most of the radiant heat will
oe stored. However, as this process of absorbing
radiant heat continues, the material becomes warmer
and less capable oE storing more heat.
is the actual cooling load that results in an avcragc
construction al)plication
with the space tempcrature held constant. The reduction in the peak
heat gain is approximately zlO(;h and the peak load
l a g s the peak heat gain by approximately 1 hour.
The cross-hatched areas (I’is. 3) represent the Heat
Stored and the Stored Heat Removed from the
construction. Since all ol the heat coming into a
space must be rcmovctl, these two areas are equal.
FIG. 3 - ACTUAL COOLING LOAD, SOLAR HEAT G AIN ,
W EST E X P O S U R E, A VERAGE C O N S T R U C T I O N
The relatively constant light load results in a
large portion being stored just after the lights are
turned on, with a decreasing amount being stored
the longer the lights are on, as illustrated in I;ig. f.
The upper and lower curves represent the instantaneous heat gain and actual cooling load from
/Z~orescent lights with a constant space temperature.
The cross-hatched areas are the Heat Stored and the ,
Stored Heat Removed from the construction. The
dotted line indicates the actual cooling load for the
first clay if the lights are on longer than the period
shown.
Figs. 3 and -f illustrate the relationship between
the instantaneous heat gain and the actual cooling
load in average construction spaces. With light construction, less heat is stored at the peak (less storage
capacity available), and with heavy construction,
more heat is stored at the peak (more storage capacity available), as shown in Fig. 5. This aspect affects
the extent of zoning required in the design of a
system for a given building; the lighter the building
construction, the more attention should be given to
zoning.
The upper curve OF I;ig. 5 is the instantaneous
solar heat gain while the three lower curves are
the actual cooling load for light, medium and heavy
constrz~tion
respectively, with a constant temperature in the space.
One more item that significantly affects the
storage of heat is the operating period of the air
conditioning equipment. All of the curves shown in
FIG. 4 - ACTUAL COOLING LOAD FROM FLUORESCENT
LIGHTS, AVERAGE CONSTRUCTION
FIG. 5 - ACTUAL COOLING LOAD, SOLAR HEAT G AIN ,
LIGHT, MEDIUM AND HEAVY CONSTRUCTION
The highly varying and relatively sharp peak of
the instantaneous solar heat gain results in a large
part of it being stored at the time oE peak solar heat
gain, as illustrated in Fig. 3.
The upper curve in Fig. 3 is typical of the solar
j~at gnin for a west exposure, and the lower curve
.
!
(:I1.\l’~l’I~,l<
3. IIL\‘I’ s’I’oli.\c;~.
I~i\‘~.IIsI’I‘\~.
.\KT)
~1‘1I.\~1’11~1~:.\
1-27
I’IOS
I;i,q.s. 3, J, ~rrlrl 5 illrlstratc the actual cooling lo;~cl
[or 2.1.horrr operxr.ion. 11’ the ccluipiicnt i s s h u t
tlotvn alter 16 li011rs ol oper;ition, some of the storctl
heat rcln;titls in tlic building construction. This
heat moist 1x2 rcniovctl @cat in miist ccl~al hcxt out)
a t i d w i l l ;ippear ;ls ;I ~~iilltlow~i
load when tlic
cquilxnci~t is turned on tlic next day, ;IS illustrated
ill Fig. 6.
:\cltiing the pidlclown loacl to the cooling load for
that day results in the actual cooling load lor
lci-hour ope~lion, as illi~stratcd in Pig. 7.
The upper curve represents the instantaneous heat
gain and the lower curve the ~C~UCLI cooling loall
for that day with a constant temperature maintained
within the space during the operating period of
the equipment. The dotted line represents the additional cooling load from the heat left in the buildconstruction. The temperature in the spacex.,..s during the shutdown period lrom the nighttime transmission load and the stored heat, and is
brought back to the control point during the pulldown period.
Shorter periods ol operation increase the pulldown load because more stored heat is left in the
building construction when the equipment is shut
off; I;ig. 8 illustrates the pulldown load for X 2 - h o u r
TIME IHR)
,
L
FIG, 8 - PULLDOWN LOAD, SOLAR HEAT G AIN ,
W I -ST EXPOSIJRE, l.Z-HOLIR O PERATION
operation.
Adding this pulldown load to the cooling load
for that day results in the actual cooling load for
12-hour operation, as illustrated in Fig. 9.
The upper and lower solid curves are the instantaneous heat gain and the actual cooling load
in average construction space with a constant temperature maintained during the operating period.
The cross-hatched areas again represent the Heat
Stored and the Stored Heat Kemoved from the
jtruction.
The light loud (fluorescent) is shown in Fig. 10 for
12- and 16-hour operation with a constant space
temperature (assuming lo-hour operation ol’ lights).
16
ACTUAL COOLING LOAD
v
TIME (HA)
12
FIG. 9 - ACTUAL COOLING LOAD, SOLAR HEAT G AIN ,
W EST EXPOSURE , 12-130~~ O PERATION
I N S T A N T A N E O U S H E A T GAP4
FIG. 6 - PULLDOWN
LOAD, SOLAR HEAT G AIN ,
W EST EXPOSURE , 16-HOIJR O PERATION
FIG. 10 - ACTUAL COOLING, LOMI FROM FLUORESCENT
LIGHTS, 12- AND I6-HOIIR O PERATION
Basis of Tables 7 thru 12
Storage Load Factors,
Fintl:
.\. The actual cooling load from the solar heat gain in July
at 4 p.m., 40” North latitude with the air conditioning
equipment operating 24 hours during the pc;t!i
loarl
pcriotls
and a constant temperature maintained within
the room.
Solar and Light Heat Gain
12., 16., and 24-hour Operation,
Constant Space Temperature
'l‘l~csc t;tl)lcs
tl~~vclopctl
These
arc
c a l c u l ; t t c t l , using
Cr01n ;I s e r i e s ol t e s t s i n
tests were contluctctl
i
n
a
actual
I)roceclurc
0. The
I)uiltlings.
office I)uildings,
su-
;~ntl residences throughout this country.
l>cr”‘;lrkcts,
J‘hc Iil;tgnitutlc
or the stor:tge effect is determined
largely by the thermal capacity or heat holding
capacity of the niatcrials surrounding the space. The
thermal capacity of a material is the weight times
the hpccific heat of the material. Since the specific
heat of most construction material is approximately
0.20 I%tu/ (lb) (F), the thermal capacity is directly
proportional to the weight of the material. Therefore, the data in the tables is based on weight of the
tcrials
surrouncling
the
space, per
cooling load at 8 p.m. for the same conditions.
Solution:
The weight per sq ft of floor arca of this room (values 01,.
tained from Chn/7ter 5) is:
Orltsitlc wall = 120 x ;; ; -<&yj x ,‘) x 1% ,l,/sq ft
(Table 31, page 64)
= 25.2 Il)/sq ft floor area
“0X8X3
x 22 Il,/sq ft
20 x 20
(Table 26, page 70)
= 13.2 lh/sq ft floor area
Partitions
= l/s x
Floor
= l/z x $zyyqf x 59 lb/q ft
i
(Table 29, page 73)
= 29.5 Ib/sq ft floor area
20 x 20
square Eoot of
110or arca.
Solar
and
Light
Heat
Gain
12-, 16-, and 24.hour Operation,
Constant Space Temperature
Tables 7 t1~~z~ II are used to determine the actual
cooling load from the solar heat gain with a constant temperature maintained within the space for
different types of construction and periods of operation. Iliith both the 12- and l&hour factors, the
starting time is assumed to be 6 a.m. suntime (7 a.m.
Daylight Saving Time). The weight per sq ft of
types of construction are listed in Tables 21 thru 33,
pages 66-76.
The actual cooling load is determined by multiplying the storage load factor from these tables
for any or all times by the peak solar heat gain for
”
e particular exposure, month and latitude desired.
:\ i’able 6 is a compilation of the peak solar heat
gains for each exposure, month and latitude. These
values are extracted from Table 15, page 44. The
peak solar heat gain is also to be multiplied by
either or both the applicable over-all factor for
shading devices (Table 16, page 52) and the corrections list& under Table 6. Reduction in solar
heat gain from the shading of the window by reveaIs
and/or overhang should also be utilized.
Example J -
20 x 20
= I/z x “0 x 59 Il,/sq ft
(Table 29,page 73)
= 29.5 Il)/sq ft floor area
Ceiling
Use of Tables 7 thrti 12
Storage Load Factors,
Actual Cooling Load, Solar Heat Gain
Given?
ti
ti
A 20 ft X 20 ft X 8 ft outside office room with &inch sand
aggregate concrete floor, with a floor tile finish, 2l/$inch
solid sand plaster partitions, no suspended ceiling, and a
12-inch common brick outside wall with i/s-inch sand aggregate plaster finish on inside surface. A 16 ft X 5 ft steel sash
window with a white Venetian blind is in the outside wall
and the wall faces west.
.
NOTE: One-half of the partition, floor and ceiling thickness is used, assuming that the spaces above and
below are conditioned and are utilizing the other
halves for storage of heat.
.
Total weight per sq ft of floor area
= 25.2 + 13.2 + 29.5 + 29.5 = 97.4 lb/sq ft.
The overall factor for the window with the white Venetian
blind is 0.56 (Table i6, page 52) and the correction for steel
sash = 1 / .85.
.I. Storage factor, 4 p.m. = 0.6G (Thble 7)
The peak solar heat gain for a west exposure in July at
40” North latitude = 164 Btu/(hr)(sq ft), (Table 6).
Actual cooling load
=
5 x 16 x 164 x .56 x-j& X 0.66 = 5700 Btu/hr
(
)
II. Storage factor, 8 p.m. = .20 (Table 7)
Actual cooling load
=
.
(
5 x 16 x 164 x .56x-$-
1
x .20 = 1730 Btu/hr
Table I-3 is used to determine the actual cooling
load from the heat gain from lights. These data may
also be used to determine the actual cooling load
from:
People - except in densely populated areas
such as auditoriums, theaters, etc. The radiant
heat exchange from the body is reduced in
situations like this because there is relatively
less surface available for the body to radiate to.
Some appliances and machines that operate
periodically, with hot exterior surfaces such
as ovens, dryers, hot tanks, etc.
NOTE: For Items 1 and 2 above, use values listed
for fluorescent exposed lights.
!
(;H:\I”I‘ER 3. I-IlC.\‘I‘ S’I‘OKAGE.
I>IVI:KSI~l‘Y,
1-X)
AI
.\Nl> ~l‘li.\‘l‘ll~l~:.\‘l’lON
= 51’31) Iltu/l1~.
ns 111~ pcoplc nrrivc ;tt R ;l.~n.).
TABLE 6-PEAK SOLAR HEAT GAIN THRU ORDINARY GLASS*
Btu/(hr)(sq it)
NORTH
LAT.
MONTH
JUne
July B May
Aug & April
Sept 8 March
Ott B Feb
Nov 8 Jon
Dee
?O
June
July K May
Aug & April
Sept 8 March
Ott S Feb
Nov B Jon
Dee
!,"+QOO
-J&l”*
July 8 May
Aug & April
Sept 8 March
Ott & Feb
Nov & Jon
Dee
20"
/-“
\
,30" )
/
‘l...-.-A’
June
July 8. May
Aug 8 April
Sept & March
Ott & Feb
Nov & Jon
D.X
JWle
July 8 May -
Aug B April
Sept 8 March
Ott & Feb
Nov 8 Jon
Dee
5 0 ”
JWle
July B May
Aug 8 April
‘Sept 8 March
Ott 8 Feb
Nov B Jon
DeC
Nt
-
59
48
25
10
10
10
10
40
30
13
10
10
9
9
'26
19
11
IO
9
8
8
20
EXPOSURE NORTH LATITUDE
NE
E
SE
156
153
141
147
152
42
52
79
118
141
153
156
14
14
I4
‘4.
34
a2
156
55
14
14
14
55
“a’
79
52
42
163
163,
163
152
147
12
148
130
103
155
158
66
155
143
137
37
28
154
,138
118
a7
52
26
18
163
164
94
127
149
106
14
14
163
73
85
113
140
26
65
147
128
121
160
164
lb7
111
141
149
161
90
100
129
152
21
30
160
163
165
164
165
11
9
a
7
158
135
116
105
6
17
15
11
9
7
5
5
133
127
102
58
35
12
10
162
lb
126
117
94
58
29
9
7
‘64
162
149
122
100
lb3
162
162
111
125
146
lb2
163
Steel Sash or
No Sash
X1/.85 or 1.17
120
63
‘05
145
159
163
54
69
102
140
162
156
a6
148
166
165
lb4
163
135
143
157
93
158
138
105
64
47
163
157
127
-116
106
‘38
158
167
153
141
E
EXPOfURE
Solar Gain
Correction
28
73
161
163
lb
14
11
8
5
4
3
5
-
66
67
Haze
-15%(Max)
42
52
79
118
141
153
66
94
127
149
101
163
73
a5
113
140
lb0
164
167
NW
ioriz
MONTH
147
156
226
Dee
Nov 8 Jon
Ott 8 Fab
Sept 8 March
Aug 8 April
July 8 May
June
I53
148
130
103
G7
37
28
250
247
230
210
202
Dee
Nov 8 Jan
Ott 8 Feb
Sept & March
Aug 8 April
July 8 May
June
154
138
118
a7
52
26
18
250
251
247
233
208
i,ao
170
DeC
Nov 8 Jan
Ott 8 F e b
Sept 8 March
Aug 8 April
July 8 May
J u n e
DeC
Nov 8 Jon
Ott 8 Feb
Sept 8 Mfarch
Aug 8 April
July & May
June
30"
12
250
246
235
212
179
145
131
133
127
102
58
35
12
10
237
233
214
183
129
103
85
Dee
Nov a?. Jan
Ott 8 Feb
Sept 8. March
Aug & April
July 8, May
June
40"
Dee
Nov & Jan
Ott & Feb
Sept 8 March
Aug & April
July & May
June
50"
152
163
lb7
163
152
147
155
158
163
lb4
155
143
137
160
163
165
163
147
128
121
I53
141
118
79
52
42
66
90
100
129
152
163
161
164
165
158
135
ye
:131,
108
90
39
162
162
116
lb
111
125
146
162
163
156
,J32,
148
135
143
157
163
157
127
SOUTH
LAT.
W
105
(64
1-6-2~
149
122
100
a6
164
lb3
158
138
105
64
116
47
N W
W
233
245
250
245
233
226
126
117
94
58
29
9
7
JO”
20"
tlorir
-
SOUTH LATITUDE
Altitude
+0.7% per 1000 ft
Dewpoint
Above 67 F
-7y0 per 10 F
Dewpoint
Below 67 F
+770 p e r 1 0 F
South Lot
Dee or Jon
+7%
*Abstracted from Table 15, page 43.
~Solar heat gain on North exposure (in North latitudes) or on South exposure (in South latitudes) consists primarily of diffuse radiation which is esrentially constant throughout the day. The solar heat gain values for this exposure ore the average for the 12 hr period (6 mm. to 6 P.?). The storage
factors in Tables 7 thru 11 (1ssutne that the solar heat gain on the North (or South) exposure is constant.
I
~ l’.\R’l I. I.O,\I) I~S’l’lhf.\-l’lN(~
l-00
TABLE 7-STORAGE LOAD FACTORS, SOLAR HEAT GAIN THRU GLASS
W&i
INTERNAL SHADE*
24 Hour Operation, Constant Space Temperaturet
EXPOSURE
(North 1.1)
WEIGHTS
(lb per sq fi
of floor area)
SUN TIME
AM
6
7
8
9
PM
10
11
12
1
2
3
4
5
6
7
A M
l
8
9
10
11
12
1
2
3
4
EXPOSURE
(South Lat)
5
Northeast
150 a over
100
30
.47 .58 54 .42 .27 .2, .20 .,P .I8 .I7 .16'.14 .I2 1.09 .08 .07 .06 '.06 .05 .05 .04 .04 .04 .03
. 4 8 . 6 0 . 5 7 . 4 6 . 3 0 . 2 4 2 0 . , 9 . , 7 ;. I 6 1. I 5 '. 1 3 . l l I . 0 8 . 0 7 . 0 6 . 0 5 . 0 5 . 0 4 . 0 4 . 0 3 . 0 3 . 0 2 . 0 2
55 .76 .73 .58 .36 .24 .,9 .17 .I5 .I3 .I2 .I1 .07 .04 .02 .02 .Ol .Ol
0
0
0
0
0
0
Southeast
East
1 5 0 8 DYB,
100
30
.39 .56 .62 .59 .49 .33 .23 2, .20'.18 ,I7 .I5 .I2 .I0 .09 .08 .08 .07 .06‘ .05 .05 .05 .04 .04
. 4 0 . 5 8 . 6 5 . 6 3 . 5 2 . 3 5 . 2 4 . 2 2 I . 2 0 . 1 8 I. 1 6 . I 4 /. I 2 . 0 9 . 0 8 . 0 7 . 0 6 . 0 5 . 0 5 . 0 4 . 0 4 . 0 3 . 0 3 . 0 2
0
0
0
0
0
0
.46 .70 .80 .79 .64 .42 .25 .,9 .,6'.14 1.11 .09 .07 .04 .02 .02 .Ol .Oi
East
150 8 eve,
100
Northeast
30
.04 .28 .47 59 .64 .62 33 .4, 27 .24 .21 .I9 .16,.14 .,2'.1, .I0 .09 .08 .07 .06 .06 .05 .05
.03 .28 .47 .61 .67 .65 .57 .44 .29 .24 .21 .I8 .I5 .I2 .lO .09 .08 .07 .06 .05 .05 .04 .04 .03
0
0
0
0
0
0
0 .30 .57 .75 .84 .8, .69 30 .30 .20 .17 .I3 .09 .05 .04 .03 .02 .Ol
150 8 OYW
100
30
. 0 6 . 0 6 . 2 3 . 3 8 5 , . 6 0 . 6 6 . 6 7 . 6 4 ' . 5 9 , . 4 2. 2 4 . 2 2 . I 9 . I 7 . I 5 . I 3 . 1 2 . I 1 . I 0 . 0 9 . 0 8 . 0 7 . 0 7
.04 .04 .22 .38 52 .63 .70 .71 .69 .59 .45 .26 .22 .I8 .I6 .13 .12 .lO .09 .08 .07 .06 .06 .05
.I0 .21 .43 .63 .77 .86 .88 .82 .56 .50 .24 .I6 .I1 .08 .05 .04 .02 .02 .Oi .Ol 0
0
0
0
North
150 a over
100
.08 .08 .09 .,O .,, .24 .39 .53 .63 .66 .61 .47 .23 .I9 .18 .16 .I4 .I3 .I1 .lO .09 .08 .08 .07
.07 .08 .08 .08 .I0 .24 .40 .55 .66 .70 .64 .50 .26 20 .I7.I5 .13 .I1 .I0 .09 .08 .07 .06 .05
Southeast
South
Southwest
30
West
Northwest
.86 .79 .60 .26 .I7 .12 .08 .05 .04 ,.03 .OZ .Ol
.Ol
0
0
150 8 ever
100
30
.08 .09 .09 .lO .lO .I0 .I0 .I8 .36 .52 .63 1.65 .55 .22 .I9 .I7 .I5 .14 .I2 .I1 .lO .09 .08 .07
.07 .08 .08 .09 .09 .09 .09 .I8 .36 .54 .6_6 .68 .60 .25 .20 .17 .I5 .13 .11 .lO .08 .07 .06 .05
.03 .04 .06 .07 .08 .08 .08 .,9 .42 .65 .81 .85 .74 .30 .I9 .I3 .OP .06 .05 .03 ,02 .02 .Ol 0
150 8 over
100
.08 .09 .lO .,O .lO .,O .,O .,O .I6 .33 .49 .61 .60 .I9 .I7 .I5 .13 .I2 .lO .09 .08 .08 .07 .06
.07 .08 .09 .09 .I0 .I0 .I0 .lO .16 .34 .52 .65 .64 23 .I8 .I5 .12 .11 .09 .08 .07 I.06 .06 .05
30
North
and
Shade
.03 .04 .06 .07 .09 .23 .47 .67 .81
150 a over
100
30
.03 .05 .07 .08 .09 .09 .I0 .I0 .17 .39 .63 .80 .79 .28 .I8 .I2 .09 .06 .04 .03 .02 .02 .Ol
.
Northwest
West
Southwest
0
.08 .37 .67 .7, .74 .76 .79 3, .83 .84 .86 .87 .88 .29 .26 .23 ,20 .I9 .17 15 .lA .I2 .ll .I0
.06 .31 .67 .72 .76 ,79 .81 .83 .85 .87 .88 .90 .91 .30 .26 .22 .19 .16 .I5 .I3 .12 .lO .09 .08
0 . 2 5 . 7 4 . 8 3 . 8 8 . 9 1 . 9 4 t.96 .96 .98 .98 .99 .99 .26 . 1 7 , . 1 2 . 0 8 . 0 5 . 0 4 . 0 3 , . 0 2 . O l , . O l . O l
South
and
Bhade
Equation: Cooling Load, Btu/hr = [Peak solar heat gain, Btu/(hr) (sq ft), (Table 6)]
x [Window oreo, sq ft]
X [Shade factor, ,Hoze factor, etc., (Chapter 4)]
X [Storage factor, (above Table at desired time)]
*Internal shading device is any type of shade located on the inside of the glass.
tThese factors apply when maintaining a CONSTANT TEMPERATURE in the space during the operating period. Where the
allowed to swing, additional storage will result during peak load periods. Refer to Table 13 for applicable storage factors.
$Weight
temperature
p e r sq ff o f floor(Weight of Outside Walls, lb) +
Room on Bldg Exterior (One or more outside walls) =
Room in Bldg Interior (No outside walls) =
Ceiling, lb)
‘% (Weight of Partitions, Floor and Ceiling, lb)
-.__
Floor Area in Room, sq ft
(Weight of Outside Walls, lb)
Basement Room (Floor on ground) = _____-__
Entire Building or Zone =
% (Weight of Partitions, Floorznd
Floor Area in Room, sq ft
+ (Weight of Floor,
lb) +___% (Weight of Partitions and
Floor Area in Room, sq ft
(Weight of Outside Wall, Partitions, Floors, Ceilings, Structural Members
ond Supports,
lb1 _____-.
-Air Conditioned Floor Area, sq ft
With rug on floor-Weight of floor should be multiplied by 0.50 to compensate for insulating effect of rug.
Weights per sq ft of common types of construction are contained in Tables 21 fhru 33, pcrges
66 thru 76.
Ceiling,~lb)
__-__
is
(:I-1.\l”l’I~:lI
3. III,:.\ I ‘
S’I 01<.\c;1:.
I)IvI~:lisl’l’\r’,
1-o 1
.\NI) s’l~I~.\‘I‘II~I~:,\‘l’loi\[
TABLE 8-STORAGE LOAD FACTORS, SOLAR HEAT GAIN THRU GLASS
WITH BARE GLASS OR WITH EXTERNAL SHADE:
24
EXPOSURE
( N o r t h Lot) 1
WEIGHT$
lib oer %a It
if fi00r Ad
Hour
Operation,
Constant
Space
Temperaturet
SUN TIME
AM
6
7,
8
9
11
12
AM
PM
I
10
1
2
3
4
6
5
7
9
8
10
1t
12
1
2
3
4
5
Northeast
150 a oY*r
100
30
.I7 27 .33 .33 21 29 .27 ,125 .23 .22 20 .,P .I7 .I5 .I4 .I2 .I1 .I0 .09 .08 .07 .07 .06 .06
.I9 .31 28 .39 .36 .34 27 .24 .22 .2l .I9 .17 .I6 .I4 .I2 .I0 .07 .08 .07 .06 .05 .05 .04 .03
.31 56 .65 .6l .46 .33 .26 .2l .I8 .I6 .I4 .I2 .09 .06 .04 1.03 .OZ .Ol .Ol .Ol
0
0
0
0
East
150 B ov*r
100
30
.I6 .26 .34 .39 .40 .38 .34 .30 .28 .26 .23 .22 .20
.I6 .29 .40 .46 .46 .42 .36 .3l .28 .25 .23 .20 .I8
.27 50 .67 .73 .68 .53 .38 .27 .22 .I8 .I5 .I2 .09
150 a OYBl
100
.08 .I4 .22 .3l .38 .43 .44 .43 I.39 .35 .32 .29'.26
.05 .I2 .23 .35 .44 .49 .51 .47 .4l .36 .3, .27 .24
Southeast
30
South
thwert,
150 & *Ye*
100
'.I1 .I0 .lO .I0 .I0 .14 .21 .29 .36 .43 .47 .46‘.40 .34 .30 .27 .24'.7.2 .20 .I8 .16 .14 .I3 .12
.09 .09 .08 .09 .09 .14 .22 .3l .42 30 .53 .5l .44 .35 .29 .26 .22 .I9 .17 .I5 .I3 .I2 .I1 .09
30
Northwest
North
/
.02 .03 .05 .06 .08 .I2 .34 53 .68 .78 .78 .68 .46 .29 .20 .I4 .09 .07 .05 .03 .02 .02 .Ol
.I2 .I1 .I1 .I0 .I0 .lO .I0 .13 .I9 .27 .36 .42 .44 .38 .33 .29 .26 23 .21 .18 .I6 .I5 .I3 .,2
.09 .09 .09 .09 .09 .09 .I0 .I2 .I9 .30 .40 .48 .51 .42 .35 .30 .25 .22 .19 .I6 .I4 .I3 .I1 .09
.02 .03 .05 .06 .07 .07 .08 .I4 .29 .49 1.67,.76 I.75 53 .33 .22 I.15 .I1
30
.02 .04 .05 .07 .08 .09 .I0 .I0 .13 I.27 .48 .65 .73 .49 .3l .21 .I6 .I0 .07 .05 .04 .03 .OZ .Ol
Equation: Cooling toad,
30
1. I 6 1. 2 3 ! . 3 3 /. 4 l 1. 4 7 /. 5 2 /. 5 7 / . 6 l /. 6 6 /. 6 9 /’7 2
i. 7 4 ’ . 5 9 i. 5 2
i
i
1. 4 6 . 4 2 1. 3 7 1. 3 4 /. 3 l /. 2 7 1. 2 5 . 2 3 / . 2 1 / . I 7 1
.I1 53 .44 .5l .57 .62 .66 .70 .74 .76 .79 .80 I.60 Sl .44 .37 .32 .29 .27 .23 .21 .18 .I6 .I3
0 .48 I.66 I.76 .82 .87 .9l I.93 .95 .97 .9e .98 I.52 .34 .24 .I6 I.11 I.07 1.05 .04 .02 .02 .Ol .Ol
Btu/hr = [Peak solar heat gain,
Btu/(hr) (sq ft),
(Table
Northwest
West
.08 .05 .04 1.03 .02 .O,
.lO .I0 .I0 .,O .I0 .I0 .I0 .I0 .I2 .I7 .25 .34 .39 .34 .29 .26 .23 .20 .I8 .16 .14 .I3 .I2 .I0
.08 .09 .09 .09 .09 .09 .09 .09 .,, .I9 I.29 .40 .46 .40 .32 .26 .22 .I9 .16 1.14 .I3 .I1 .I0 .08
150180~Ver
North
.O,
150 8 OVW
100
and
Shade
East
Northeast
.I0 .I0 .13 .20 .28 .3S .42 .48 .5l .5l ,413 .42 .37 .33 .29 .26 .23 .2l .I9 .I7 .I5 .I4 .I3 .12
.07 .06 .I2 .20 .30 .39 .48 .S4 .58 .57 .53 .45 .37 .3l .27 .23 .20 .I8 .I6 .I4 .I2 .I1 .I0 .08
0
0 .I2 .29 .48 .64 .75 .82 .8l .75 .6l .42 ,213 .I9 .I3 .09 .06 .04 .03 .02 .Ol .Ol
0
0
150 8 OYW
100
Southeast
.I8 .40 .59 .72 .77 .72 .60 .44 .32 .23 .I8 .14
150 a over
100
30
30
West
0
.I3 .I2 .lO .09 .08 .OB .07 .06
.I1 .09 .08 .08 .06 .06 .05 .04
.02 .Ol .Ol .Ol .Ol
0
0 .Ol
EXPOSURE
(South L-t)
Southwest
I
South
and
Shade
611
X
[Window area, sq ft]
x [Shade factor, Haze factor, etc., (Chapter 4)]
X [Storage factor, (above Table at desired time)]
$Bare glass-Any window with no inside shading device. Windows with shading devices on the outside or shaded by external projections are
considered bare glass.
tThese factors apply when maintaining a CONSTANT TEMPERATURE in the space during the operating period. Where the
allowed to swing, additional storage will result during peak load periods. Refer to Table 13 for applicable storage factors.
$Weight
temperature
per sq ft of floor-
,om on Bldg Exterior (One or more outside walls)
Room in Bldg interior (No outside walls) =
(Weight of Outside Walls, lb) + % (Weight of Partitions, Floor and Ceiling, lb)
= --.-Floor Area in Room, sq ft
‘% (Weight of Partitions, Floor and Ceiling, lb)
~~~~~- --- -.-----v--.
- Floor Area m Room, sq ft
(Weight of Outside Walls, lb) + (Weight of Floor,lb)+ %-(Weight
Basement Room (Floor on ground) = ~----~
Floor Area in Room, sq ft
of Partitions and Ceiling, lb)
(Weight of Outside Wall, Partitions, Floors, Ceilings, Structural Members and Supports,lb)
Entire Building or Zone =
Air Conditioned Floor Area, sq ft
With rug on floor-Weight of floor should be multiplied by 0.50 to compensate for insulating effect of rug.
Weights per sq ft of common types of construction ore contained in Tables
21 fhru 33, pages 66 fhru 76.
is
I l-02
,
I’;\R’I‘ I . LO,\I) I~S’I~IIi.\
I’IN(i
TABLE 9-STORAGE LOAD FACTORS, SOLAR HEAT GAIN THRU GLASS
WITH INTERNAL SHADING DEVICE*
16 Hour Operation, Constant Space Temperaturet
EXPOSURE
(North Lot)
SUN TIME
WElGHlg
(lb Per rq fl
of floor area)
EXPOSURE
(South Lot)
Northeast
150 8 a”*r
100
30
Southewt
East
150 & OYB,
100
30
E0tt
Southeast
1 5 0 a OVB,
100
30
Northeast
South
150 a oy~r
100
30
North
Southwest
150 8 over
100
30
20
.08
.I9
.OB
.18
.09
.I7
.09
.18
.I0
.31
.24
.46
.47
.60
.67
.66
.a1
.70
.a6
.64
.79
.50
/.60
.26
.26
.20
.I7
.I7
.I2
.I5
.08
Wart
150 & ov*r
100
30
.23
.22
.12
.23
.21
.lO
.21
.I9
.10
.21
.I9
.I0
.20
.I7
.lO
.I9
.16
.I0
.18
.15
.09
.25
.23
.19
.36
.36
.42
.52
.54
.65
.63
.66
.65
.68
.a5
.55
.60
.74
.22
.25
.30
.I9
.20
.I9
.17
.17
.13
150 8 OVB,
100
30
.21
.19
.12
.21
.I9
.I I
.20
.I8
.I1
.I9
.I7
.1 I
.ia
.ia
.I6
.11
.I7
.I6
.ll
.16
.15
.lO
.I6
.I6
.I7
.33
.34
.39
.49
.52
.63
.61
.65
.a0
.60
.23
.79
.19
.I7
.l I
.28
.I7
.I5
.18
.I5
.12
.I2
-.* 150 a over
100
30
.23
.25
.07
.5a
.46
.22
.75
.73
.69
.79
.7a
.a0
30
.a2
.a6
.a0
.a1
.a2
.93
.a3
.94
.a2
.a4
.95
.a3
.a5
.97
.a4
.a7
.90
.86
.a8
.9a
.a7
.a9
.99
.aa
.90
.99
.39
.40
.35
.35
34
.23
.31
.29
.16
Northwest
North
ond
Shade
Equation: Cooling
.
.a1
.Ia
Northwest
West
Southwest
South
, and
Shade
Load, Btu/hr = (Peak solar heat gain, Btu/(hr) (rq ft), (Table 6)]
x [Window area, sq ft]
x [Shade factor, Haze factor, etc., (Chapter 411
x [Storage factor, (above Table at desired time)]i
*Internal shading device is ony type of shade located on the inside of the glass.
tThese factors apply when maintaining a CONSTANT TEMPERATURE in the space
allowed to swing, additional storage will result during peak load periods. Refer to
$Weight
during
Table
the operating period. Where the
13 for applicable storage factors.
temperature
p e r rq ft o f floor(Weight of Outside Walls, lb) + % (Weight of Partitions, Floor and Ceiling, lb)
Room on Bldg Exterior (One or more outside walls)
=
Floor Area in Room, sq ft
I,$ (Weight of Partitions, Floor and Ceiling, lb)
Room in Bldg interior (No outside
walls)
=
Floor Area in Room, sq ft
(Weight of Outside Walls, lb) + (Weight of Floor,
Basement Room [Floor on ground) =
Entire Building or Zone =
lb) + % (Weight of Partitions and Ceiling, lb)
__--. .-
Floor Area in Room, sq ft
(Weight of Outside Wall, Partitions, Floors, Ceilings, Structural Members and Supportd
Air Conditioned Floor Area, sq ft
With rug on floor-Weight of floor should be multiplied by 0.50 to compensate for insulating effect of rug.
Weights per sq ft of common types of construction ore contained in Tables 21 thru 33, pages 66 thru
.
76.
is
TABLE IO-STORAGE LOAD FACTORS, SOLAR HEAT GAIN THRU GLASS
W I T H B A R E G L A S S O R W I T H E X T E R N A L SHADE$
16 Hour Operation, Constant Space Temperature?
EXPOSURE
(North 1.1)
SUN TIME
WElGHT$
( l b persq ft
of floor oreal
A M
6
7
8
9
Northeast
150 8 OVB,
100
30
.28 / .37 / .42 / .41
.20
.39
33
.57
East
150 8 OYW
100
30
.29
.27
.29
.3a
.3a
.51
Southeast
150 8 OVB,
100
30
.24
.I9
.03
South
150 a DYBl
100
30
Southwest
150 a OVW
100
30
PM
10
11
1 . 3 a / .36
12
1
2
.31
1 .33
3
4
5
6
9
I .I2
.I0
.03
Southeast
.22
.20
.I2
.20
.I8
.09
.I8
.I5
.06
.I6
.I4
.04
.I4
.I2
.03
.29
.24
.20
.32
.31
.23
.29
.27
.I8
.26
.24
.I4
.23
.21
.09
.21
.ltl
.07
.19
.I6
.05
Northeast
.33
.27
.Ob
.31
.24
.04
.46
.53
.bl
.42
.45
.42
.37
.37
.2a
.33
.31
.I9
.29
.27
.I3
.26
.23
.09
North
.35
.31
.ll
.32
.2a
.lO
.09
.34
.44
.46
.30
.35
.29
.27
.29
.20
.24
.26
.14
Northwest
.I0
.40
.51
.ba
.46
.48
.53
.41
.4l
.3a
.I0
.2a
.26
.I4
.30
.33
.35
.54
.ba
.46
1 .53?
.78
.7a
West
150 8 OVW
100
30
.3a
.34
.I7
.34
.31
.I4
.32
.20
.I3
.2a
.25
.I 1
.26
.23
.l I
.25
.22
.lO
.23
.21
.I0
.25
.21
.I5
.26
.23
.29
.27
.30
.a9
.36
.40
.67
.42
.48
.76
.44
.5l
.75
Northwest
150 a ov*r
100
30
.33
.30
.lS
.30
.2a
.14
.28
.25
.I2
.26
.23
.I2
.24
.22
.12
.23
.20
.12
.22
.19
.I2
.20
.I7
.ll
.lS
.17
.I3
.I7
.lP
.27
.25
.29
.4a
.34
.40
.65
.39
.46
.73
North
and
Shade
150 8 OYer
100
30
.31
.30
.04
.57
.47
.07
.64
.ba
.72
.73
.73
.74
.74
.75
.76
.7a
.a2
.90
Equation: Cooling Load,
6
.23
.23
.I5
t
.
7
I . 2 3 I . 2 2 I 2 0 I . I 9 I . I 7 I .15 I .I4
.I9
.I7
.I6
.I4
.I2
.14
.I2
.09
.Ob
.04
EXPOSURE
(South Lal)
.a1
+
.97
Btu/hr = [Peak solar heat gain, Btu/(hr) (sq ft), (Table
?
Southwert
’
South
and
Shade
6)]
X [Window area, sq ft]
x [Shade factor, Haze factor, etc., (Chapter 4)]
X [Storage factor, (above Table at desired time)]
*Bare glass - Any window with no inside shading device. Windows with shading devices on the outside or shaded by external projections ore
considered bare glass.
?These
factors apply when maintaining CI CONSTANT TEMPERATURE in the space during the operating period. Where the
allowed to swing, additional storage will result during peak load periods. Refer to Table 13 for applicable storage factors.
$Weight
‘.
temperature
p e r sq ft o f f l o o r -
kcxxn on Bldg Exterior (One or more outside walls)
(Weight of Outside Walls, lb) + __% (Weight of Partitions, Floor and Ceiling, lb)
= __
Floor Area in Room, sq ft
% (Weight of Partitions, Floor and Ceiling, lb)
Room in Bldg interior (No outside walls) = --Floor Area in Room, sq ft
(Weight of Outside Walls, lb)
Basement Room (Floor on ground) = ~
Entire Building or Zone =
+ (Weight of Floor, lb)
+ ‘/2 (Weight of Partitions -.
and Ceiling, lb)
Floor Area in Room, sq ft
(Weight of Outride Wall, Partitions, Floors, Ceilings, Structural __.-..
Members and Supports, 5
Air Conditioned Floor Area, sq ft
With rug on floor-Weight of floor should be multiplied by 0.50 to compensate for insulating effect of rug.
Weights per sq ft of common types of construction are contained in Tables 21
thru 33, pages 66 thru
76.
is
TABLE II-STORAGE LOAD FACTORS, SOLAR HEAT GAIN THRU GLASS
12 Hour Operation, Constant Space Temperature1
INTERNAL
EXPOSURE
(North La0
WEIGHT5
(lb per tq ft
of floor area)
BARE GLASS OR EXTERNAL SHADE1
SHADE*
SUN TIME
PM
AM
6
1
B
r9
41
11
12
1
2
3
AM
4
5
6
7
a
EXPOSURE
(South Lat)
PM
9
IO
11
12
1
2
3
4
5
Northearl
150 a ov*r
100
30
. 4 2 . 4 7 I . 4 5 I. 4 2 I . 3 9 I . 3 6 . 3 3 1 . 3 0 . 2 9 ' . 2 6 . 2 5
.45 SO .49 ..45 .42 ..34 .30 .27 .26 .23 .20
.62 .69 .64 .48 .34 .27 .22 .I8 .16 .I4 .I2
Eort
150 a oYw
100
30
.51 .66 .71 .67 .57 .40 .29 .26 .25 .23 .2l .I9 .36 .44 SO .53 .53 SO .44 .39 .36 .34 .30 .28
.52 .67 .u .70 .58 .40 .29 .26 .24 .21 .I9 .16 .3P .44 .54 .58 .57 .51 .44 .39 .34 .31 .28 .24
. 5 3 . 7 4 . a 2 .a1 .65 .43 .25 .I9 .I6 .I4 . I 1 . 0 9 . 3 6 . 5 6 . 7 1 . 7 6 . 7 0 S 4 . 3 9 . 2 8 . 2 3 . I 8 . I 5 . I 2
Earl
Southeasl
150 8 ovw
100
30
.20 .42 .59 .70 .74 .71 .61 .48 .33/.30' .26 .24 .34 .37 .43 I.50 .54 .58 .57/.55 .50 .45 .4l .37
.18 .40 .57 .70 .75 .72 .63 .49 .34 .28 .25 .21 .29 .33 .41 .51 .58 .61 .61 .56 .49 .44 .37 .33
.09 .35 .61 .78 236 .a2 .69 .50 .30 .20 .I7 .I3 .I4 .27 .47 .64 .75 .79 .73 .61 .45 .32 .23 .lf
Northeast
South
150 a OVB,
100
30
. 2 8 . 2 5 . 4 4 . 5 3 .b4 .72 .77 77 .73 .67 .49 .31 .47 .43 .42 .46 .51 .56 .61 .65 .66 .65 .bl .54
.26 .22 .3a .51 .64 .73 .79 .79 .77 .65 .51 .31 .44 .37 .39 .43 SO .57 .64 .68 .70 .68 .63 S?
.21 .29 .48 .67 .79 .a8 .a9 .a3 .56 .50 .24 .I6 .2a .I9 .25 .38 .54 .68 .78 .a4 .a2 .76 .61 .4i
150 a OVBI
100
.31 .27 .27 .26 .25 .27 SO .63 .72 .74 .69 .54 Sl .44 .40 .37 .34 .36 .41 .47 .54 .57 .60 .5f
.33 .28 .25 .23 .23 .35 .50 .64 .74 .77 .70 .55 .53 .44 .37 .35 31 .33 .39 .46 .55 .62 .64 .6(
.29 .21 .18 .I5 .14 .27 .50 .69 .a2 .87 .79 .60 .48 .32 .25 .20 .I7 .I9 .39 .56 .70 30 .79 .6(
Southwest
.44 .39 .36 .33 .31 .31 .35 .42 .49 .5
.44 .39 .34 .31 .29 .28 .33 .43 .51 .5i
.38 .28 .22 .I8 I.16 .19 .33 .52 .69 .7;
Weat
Northwest
30
North
ond
Shade
150 8 over
100
30
Equation: Cooling Load,
.39 .36 33 .30 .28 .26 .26 .30 .37 .4d
.41 .35 I.31 .28 .25 23 .24 .30 .39 .41
22 .33 .25 .20 ,113 .I5 .14 .13 .I9 .41 .64 .80 .75 .53 .36 28 .24 .19'.17 .I5 .I7 .30 50 .6(
. 9 6 . 9 6 . 9 6 . 9 6 /. 9 6 I . 9 6 I. 9 6 I . 9 6 I . 9 6 I. 9 6 1 . 9 6 I. 9 6 . 7 5 . 7 5 . 7 9 2 3 3 . 8 4 . 8 6 . 8 8 3 8 1 . 9 1 . 9 2 . 9 3 . 9 :
. 9 8 .9a .9a .98 .9a .9a .9a I .9a .9a .9a I .9a .9a .a1 .a4 .a6 .a9 .91 .93 . 9 3 . 9 4 . 9 4 . 9 5 . 9 5 . 9 :
4
I . o o - - - - +b - -
__-
I
.oo
-----
Southeast
North
.
Northwest
West
Soulhwert
South
and
Shade
Btu/hr = [Peak solar heat gain, Btu/(hr) (sq ftt), (Table 6)]
x [Window area, sq ft]
x [Shade factor, Hare factor, etc., (Chapter 411
x [Storage factor, (above Table at desired time)]
*Internal shading device is any type of shade located on the inside of the glass.
$Bore glass-Any window with no inside shading device. Windows with shading devices on the outside or shaded by external projections are to
considered bare gloss.
tThese factors apply when maintaining a CONSTANT TEMPERATURE in the space during
allowed to swing, additional storage will result during peak load periods. Refer to Table
SWeight
the operating period. Where the
13 for applicable storage factors.
temperature
p e r sq f t o f f l o o r -
Room on Bidg Exterior (One or more outside walls) =
Room in Bldg Interior (No outside walls) =
Basement Room (Floor on ground) =
Entire Building or Zone =
1%
(Weight of Outside Walls,
lb) + % (Weight of Partitions, Floor .__and Ceiling, lb)
__-___..
Floor Area in Room, sq ft
(Weight of Partitions, Floor and __Ceiling, lb1
Floor Area in Room, sq ft
(Weight of Outside Walls, lb) + (Weight of Floor, lb) + % (Weight of Partitions and Ceiling, lb)
Floor Area in Room, sq ft
(Weight of Outside Wall, Partitions, Floors, Ceilings, Structural Members and Supports, J-j
Air Conditioned Floor Area,
sq ft
With rug on floor-Weight of floor should be multiplied by 0.50 to compensate for insulating effect of rug.
Weights per sq ft of common types of construction are contained in TabJes 21 thru 33, pages 66 thru 76.
is
,ed
(:tl.\l”l‘l-l~
3. IHI<.\’
s’I‘oII,\c;I:,
I~lVEI<Sl’I’Y.
.\NI)
l-35
S’II~.\‘I‘II~l(:.\‘I‘IoN
i”‘L /
TABLE 12-STORAGE LOAD FACTORS, HEAT GAIN-LIGHTS*
Lights
On
EQUIP.
OPERATION
Hour5
10
Hours-1
with
Equipment
WEIGHT8
( l b per ~4 f t
of floor orea)
0
/
150 8 o v e r
Operating
12,
NUMBER
OF
16
and
HOURS
24
Hours,
AFTER
Constant
LIGHTS
ARE
Space
TURNED
i
Temperature
r
ON
f
1
2
3
4
5
6
7
8
9
10 11
12 13 14 15 16 17 18 19 20 21
i
/
I
i
I
!
I
I
/
,
,
,
,
,
,
1
1.371.671 . 7 1 (.741.76j.79/.81/.83(.84/.861.87j.291.26 1.23/.20/.19/.17/.151.141.121.11I.10 . I
12 .I1
0 9
.08
0
0
“These factors apply when maintaining CI CONSTANT TEMPERATURE in the space during the operating period. Where the temperature 1s allowed t0
swing, odditionai storage will result during peak load periods. Refer to Table 13 for applicable storage factors.
With lights operating the scltne number of hours os the time of equipment operation, use a load factor of .l.OO.
tLights
On for Shorter or Longer Period than 10 Hours
3. Equipment operating for 12 hours:
rlccosionally adjustments may be required to take account of lights
ating less or more than the IO hours on which the table is based.
I&~= following is the procedure to adjust the load factors:
Follow procedure in Step 2, except in Step 2b add values of
12th hour to that designated 0, 13th hour to the 1 st hour, etc.
A - W I T H L I G H T S I N O P E R A T I O N F O R S H O R T E R P E R I O D T H A N IO
HOURS and the equipment operating 12, 16 or 24 hours at the time
of the overall peak load, extrapolate load factors as follows:
B-WITH LIGHTS IN OPERATION FOR LONGER PERIOD THAN 10
HOURS and the equipment operating 12, 16 or 24 hours at the time
of the overall peak load, extrapolate load factors OS follows:
I. Equipment operating for 24
hourr:
I. Equipment operating for 24 hours:
o. Use the storage load factors as listed up to the time the lights
care turned off.
a. Use the load factors as listed through 10th hour and extrapolate
beyond the 10th hour at the rate of the last 4 hours.
b. Shift the load factors beyond the 10th hour (on the right of
heavy line) to the left to the hour the lights are turned off.
This leaves last few hours of equipment operation without
designated load factors.
b. Follow the same procedure as in Step lb of “A” except shift
load factors beyond 10th hour now to the right, dropping off
the lost few hours.
c. Extrapolate the last few hours at the same rate of reduction
as the end hours in the table.
2. Equipment operating for 16 hours:
a. Follow the procedure in Step I, using the storage load factor
v(~Iues in 24-hour equipment operation table.
b. Now construct a new set of load factors by adding the new
values for the 16th hour to that denoted 0, 17th hour to the
I st hour, etc.
C. The load factors for the hours succeeding the switching-off the
lights ore as in Steps 1 b and I c.
2. Equipment operating for 16 hours or 12 hours:
a. Use the load factors in 24-hour equipment operation table 0s
listed through 10th hour and extrapolate beyond the 10th
hour at the rate of the last 4 hours.
b. Follow the procedure in Step I b of “A” except shift the load
factors beyond 10th hour now to the right.
C.
For Ibhour equipment operation, follow the procedure in
Steps 2b and 2c of “A”.
d. For I?-hour
equipment operation, follow the procedure in
Step 3 of “A”.
/
t
)\
1-36
, I’:\RT I. LOAD ESTIM:\TING
Example
Adjust values for 24-hour equipment operation and derive new values for lb-hour equipment operation for fluorescent lights in operation 8 and
13 hours, and an enclosure of 150 Ib/sq ft of ftoor.
EOUIP
WEIGHT$
O P E R A T I O N ( l b p e r sq f t
Hours
of floor area)
24
16
150
150
NUMBER OF HOURS AFTER LIGHTS ARE TURNED ON
19 120
21
22
23
LIGHTS
N
Hours
.17
.I5
.14
.12
.ll
13
.lO
.09
.OB .07
.06
.I2 .I1 ,.I0 .09
.08
8
10
O
1
3
.37
.37
.37
.67 .71
.b7 .71
.b7 .71
.74
.76
.74 .76
.74 .76
.79 .81 .a3 .84
.79 .81 .83 .84
.79 31 .83 “84
.86
.87 .89
.90 392
.29 .26 .23
.20 .19
.29
,215 .23
.20
.17
.86
.87 29 .26 .23
.I1
.I5 .I4
.bO
.87 .90
.91 .91
.93 .93
.94 .94
.95
.51
.79
.84
.87
.88
.89
.29
.84 .84
.85
.85 .86
.60 .82
.82
.83 .84
4
.85
5
6
9
2
0
7
8
.90
10
11
12
13
.19
14
15
16
17
18
.I5 .I4 .12
.20 .19 .I7
.95 .96 .96 .97 .29 .26
y .23 .20 .19 .17 .I5
.88 %O \.32 .28 .25 .23 .19
13
6
10
$Weight per rq ft of floorRoom on Bldg Exterior (One or more outside walls) =
Room in Bldg Interior (No outride walls).=
Basement Room (Floor on ground)
Entire Building or Zone =
(Weight of Outside
lb)
___--- - Walls,
- - - lb) + % (Weight of Partitions, Floor and Ceiling,
__Floor Area in Room, sq ft
(Weight of Outside Walls, lb) + (Weight of Floor, lb) + %-.__
(Weight --__
of Partitions and Ceiling, lb)
=
Floor Area in Room, sq ft
(Weight of Outside
Partitions, Floors, Ceilings, Structural Members and Supports. lb)
-~~. Wall
.A--.
Air Conditioned Floor Area, sq ft
With rug on floor-Weight of floor should be multiplied by 0.50 to compensate for insulating effect of rug.
Weights per sq ft of common types of construction ore contained in Tables 21 thru 33, pages 66 thru 76.
SPACE
TEMPERATURE
t
l/z (Weight of Partitions, Floor and
- Ceiling, lb)__Floor Area in Room, sq ft
SWING
In addition to the storage of radiant heat with a
constant room temperature, heat is stored in the
building structure when the space temperature is
forced to swing. If the cooling capacity supplied to
the space matches the cooling load, the temperature
in the space remains constant throughout the
operating period. On the other hand, if the cooling
capacity supplied to the space is lower than the
actual cooling load at any point, the temperature
in the space will rise. As the space temperature
increases, less heat is convected from the surface
and more radiant heat is stored in the structure.
This process of storing additional heat is illustrated
in Fig. 11.
The solid curve is the actua1 cooling load from
the solar heat gain on a west exposure with a constant space temperature, 24-hour operation. Assume
that the maximum cooling capacity available is represented by A, and that the capacity is controlled to
maintain a constant temperature at partial load.
When the actual cooling load exceeds the available
cooling capacity, the temperature will swing as
shown in the lower curve. The actual cooling load
with temperature swing is shown by the dotted
line. This operates in a similar manner with
different periods of operation and with different
types of construction.
NOTE: When a system is designed for a temperature swing,.
the maximum swing occurs only at the peak on design
clays, which are defined as those days when all loads
simultaneously peak. Under normal operating conditions, the temperature remains constant or close to
constant.
Basis of Table 13
- Storage Factors,
Space Temperature Swing
The storage factors in Table 13 were computed
using essentially the same procedure as Tables 7
thru 12 with the exception that the equipment
capacity available was limited and the swing in
room temperature computed.
FIG. 11 - ,~CTUAL COOLING L OAD W ITH V ARYING
R OOM T E M P E R A T U R E
The magnitude of the storage effect is determined
largely by the thermal capacity or heat holding
*
;
CH/\I”J‘JCJ<
: I . I-JJ:.\‘J‘
.S~J‘(>R;\GL’:,
J)l\il:JLSI’J‘Y,
1-37
,\ Xl> S’J‘JI,\-J‘II~I~:,\‘J‘ION
TABLE 13-STORAGE FACTORS, SPACE TEMPERATURE SWING
Btu/(hr) (deg F swing) (sq ft of floor area)
NOTE: This reduction is to be taken at the fime of peak load only.
TYPE APPLICATION
Load Pattern
I
A
VARIABLE
24-HOUR
-‘dldg
150
and
OV.3
Bidg
Periphery,
Except
North Side
30
30
I
CONSTANT
INTERMITTENT
P4-HOUR
PERIOD
Apartment
g+\
VARIABLE
24-HOUR
CONTINUOUS
PERIOD
HOUS.3,
Hotels,
Hospitals
Residences
(%I
I
75
50
25
75
50
25
100
I
Interior
zonert
Department
storer,
Factories
GLASS
RATIOS
floor
area)
Type
Office
INTERMITTENT
PERIOD
HOURS
WEIGHT
(Ib/rq f t
I
75
50
25
I
OF
OPERATION
16
1
12
Temperature Swing (F)
1-2 13-4 ( 5 - 6 ( l-2 1 3 - 4 ( 5 - 6 ( 1-2 ( 3 - 4 / 5 - 6
24
1.90 1 1 . 8 0 / 1.65 1 1.80 1 1.70 / 1.55 1 1.60 / 1.50 / 1.40
I 1.70 I 1.60 I 1.45 I 1.60 I 1.50 I 1.35 I 1.50 I 1.35 I 1.25
1.50
1.40
1.40
1.30
1.30
1.20
1.70
1.60
1.45
1.50
1.45
1.35
1.40
1.35
1.30
1.50
1.40
1.30
1.35
1.30
1.20
1.30
1.25
1.10
1.35
1.25
1.20
1.25
1.00
.90
1.20
.95
.70
1.20
1.10
.95
1.00
.95
1.40
1.25
1.00
.88
1.20
.95
.80 1.10
.90
.80
.90
.85
.80
.90 j .80
.70 / .85 / .75 / .60 j 30 / .70 1 .55
I
I
I
I
I
I
I
I
I
150 and
OVW
100
30
-
1.60
1.40
.95
1.55
1.38
.92
1.50
1.36
.90
1.50
1.30
.90
1.45
1.28
.88
1.25
.85
1.35
1.25
.a5
1.20
.80
-
150
and
Over
75
50
25
1.85
1.65
1.45
1.75
1.50
-
1.40
-
-
_
_
-
-
-
75
50
25
75
50
25
1.55
1.40
1.30
1.20
1.10
.85
1.45
1.35
1.10
.90
.70
1.40
.95
_
_
-
_
-
-
-
100
30
30
-
Equation: Reduction in Peak Cooling Load, Btu/hr = (Floor Area, sq ft) X (Desired Temp Swing, Table 4, page 20) X (Storage Factor, above table)
*Weight per sq ft of floor may be obtained .from equation on pc~ge 30
iFor 12-hour operation, use a 2 degree max temp swing.
$Glars ratio is the percent of glass area to the total wall area.
capacity of the materials surrounding the space. It is
limited by the amount of heat available for storage.
Load patterns for different applications vary approximately as shown in the first column of TnOie 1.3.
For instance, an office building has a rather large
varying load with a high peak that occurs intermittently. An interior zone has an intermittent peak
the load pattern is relatively constant. A hospital, on the other hand, has a constant base load
which is present for 24 hours with an additional
intermittent load occurring during daylight hours.
The thermal capacity of a material is the weight
times the specific heat of the material. Since the
specific heat of most construction material is approximately 0.20 Btu/(lb)(F), the thermal capacity
is directly proportional to the weight of the material.
Therefore, the data in the tables is based on weight
of the materials surrounding the space, per square
foot of floor area.
Use of Table 13
- Storage Factors,
Space Temperature Swing
.
-5
Table 23 is used to determine the reduction in
cooling load when the space temperature is forced
to swing by reducing the equipment capacity below
that required to maintain the temperature constant.
This reduction is to be subtracted from the room
sensible heat.
NOTE: This reduction is only taken at the time of
cooling
load.
Example 3 - Space Temperature Swing
Given:
Find:
The actual cooling load at 4 p.m. from sun. lights, and
people with 3 F temperature swing in the space.
Solution:
The peak sensible cooling load in this room from the sun,
I igh ts, and people (neglecting transmission inliltration.
ventilation and other internal heat gain) is
5700 + 5190 = 10,590 Btu/hr.
(Esavtf11es 1 and 2.)
NOTE: The peak cooling load in this room occurs at approximately 4 p.m. The solar and light loads are
almost at their peak at 4 p.m. Although the transmission across the large glass window peaks at
about 3 p,m., the peak infiltration and ventilation
load also occurs at 3 p.m. and the relatively small
transmission load across the wall peaks much later
at about 12 midnight. The sum of these loads re-
’
Since the nortual
t h e
tlwrtnostat
scttina i s ;~i)out 7 5 1: o r iti I: (II),
tlcsijin temlxrattlr~ (78 1: = 73 T: thcrnwstat setting
+ :I F s w i n g ) OCCIII-s o n l y on tlcsigrl lxxk days a t t h e t i m e
o[ peak
loacl. Untler partial load
olxzl.ation,
tllc r00l11
perature is ktwcen 75 I: dl) a n d i8 I; tll), o r a t the
tCI11-
thertnostat
s e t t i n g (7.5 1:) , tlclwn~lin~ on the loatl.
PRECOOLING
STORAGE
AS A MEANS OF INCREASING
Precooling a space below the temperature norinally desired illcl-eases the stoqe of 1~eat a t t h e
time of peak load, only when the precooling temperature is maintained as the control point. This
is because the potential temperature swing is increased, thus adding to the amount of heat stored
at the time of peak load. Where the space is precooled to a lower temperature and the control
point is reset upward to a comfortable condition
when the occupants arrive, no additional storage
occurs. In this situation, the cooling unit shuts off
and there is no cooling during the period of warming up. When the cooling unit begins to supply
cooling again, the cooling load is approximately
up to the point it would have been without any
precooling.
Precooling is very useful in reducing the cooling
load in applications such as churches, supermarkets,
theaters, etc., where the precooled temperature can
be maintained as the control point and the temperature swing increased to 8 F or 10 F.
DIVERSITY OF COOLING LOADS
Diversity of cooling load results from the probable
non-occurrence of part of the cooling load on a
design day. Diversity factors are applied to the
refrigeration capacity in large air conditioning systems. These factors vary with location, type and
size of the application, and are based entirely on
the judgment of the engineer.
Generally, diversity factors can be applied to
people and light loads in large multi-story office,
hotel or apartment buildings. The possibility of
having all of the people present in the building
and all of the lights operating at the time of peak
load are slight. Normally, in large office buildings,
sonic people will be away from the office on other
I,usincss. i\lso, the lighting arrangement will freclucntly 1x2 such that the lights in the vacant ofkes
will not be on. In addition to lights being elf becaiisc the people are not present, the normal maintenance procedure in large office buildings usually
rest110 in some lights being inoperative. Therefore,
a diversity Iactor on the people a n d light loads
shoultl
be applied for selecting the proper size
refrigeration ecluilment,
The silt of the diversity factor depends on the
size of the I~uiltling and the engineer’s judgment
of the circumstances involved. For example, the
diversity factor on a single small office with 1 or 2
people is 1.0 or no reduction. Expanding this to
one iloor of a building with 50 to 100 people, 5y0
to lO(70 may be absent at the time of peak load,
and expanding to a 20, SO or 40-story building, 10% ’
to 207’ may be absent during the peak. A building
with predominantly sales offices would have many
people out in the normal course of business.
This same concept applies to apartments and
hotels. Normally, very few people are present at the
time the solar and transmission loads are peaking,
and the lights are normally turned on only after
sundown. Therefore, in apartments and ho$els, the
diversity factor can be much greater than with office
building!.
These reductions in cooling load are real and
should be made where applicable. Table 14 lists
some typical diversity factors, based on judgment
and experience.
TABLE I$-TYPICAL DIVERSITY FACTORS
FOR LARGE BUILDINGS
(Apply to Refrigeration Capacity)
DIVERSITY FACTOR
TYPE OF
APPLICATION
Office
Apartment,
Department
Hotel
Store
Industrial*
People
Lights
.75 to .90
.70 to .05
.40 to .60
.30 to .50
.BO to .90
.90 to 1.0
.85 to .95
.BO to .90
Equation:
Cooling Load (for people and lights), Btu/hr
= (Heat Gain, Btu/hr, Chapter 7)
X (Storage Factor, Table 121% (Diversity Factor, above table)
*A diversity factor should
Refer to Chopfer 7.
also
be
applied
to
the
machinery
load.
Use of Table 14
- Typical Diversity Factors for Large Buildings
The diversity factors listed in Table 14 are to be
used as a guide in determining a diversity factor
for any particular application. The final factor must
necessarily be basctl on judgment of the effect of the
m a n y variables irlvolvetl.
STRATIFICATION OF HEAT
\
There are generally two situations where heat is
stratified and will reduce the cooling load on the
air conditioning equipment:
1. Heat may bc stratified in rooms with high
ceilings where air is exhausted through the
root’ or ceiling.
2. Heat may be contained above suspended ceilings with rccessecl lighting and/or ceiling
plenum return systems.
The first situation generally applies to industrial
applications, churches, auditoriums, and the like.
The second situation applies to applications such
qflice buildings, hotels, and apartments. With
t,uch cases, the basic fact that hot air tends to rise
makes it possible to stratify loads such as convection
from the root’, convection from lights, and convection from the upper part of the walls. The convective portion of the roof load is about 25% (the
rest is radiation); the light load is about 50% with
fluorescent (20% with incandescent), and the wall
transmission load about 40’%.
In any room with a high ceiling, a large part of
the convection load being released above the supply
air stream will stratify at the ceiling or roof level.
Some will be induced into the supply air stream.
Normally, about 80% is stratified and 20y0 induced
in the supply air. If air is exhausted through the
ceiling or roof, this convection load released above
the supply air ~riay IX subtracted from the air conditioning load. This results in a large reduction
in load if the air is to be exhausted. It is not normally practical to exhaust more air than necessary,
as it must be made u p by bringing outdoor air
through the apparatus. This usually results in a
larger increase in load than the reduction realized
by exhausting air.
Nominally, about a 10 F to 20 F rise in exhaust air
temperature may be figured as load reduction if
there is enough heat released by convection above
the supply air stream.
Hot air stratifies at the ceiling even with no
exhaust but rapidly builds up in temperature,
and no reduction in load should be taken where
air is not exhausted through the ceiling or roof.
With suspended ceilings, some of the convective
heat from recessed lights flows into the plenum
space, Also, the radiant heat within the room (sun,
lights, people, etc.) striking the ceiling warms it
up and causes heat to flow into the plenum space.
These sources of heat increase the temperature of
air in the plenum space which causes heat to flow,
into the underside of the floor structure above.
When the ceiling plenum is used as a return air
system, some of the return air flows through and
over the light fixture, carrying more of the convective heat into the plenum space.
Containing heat within the ceiling plenum space
tends to “flatten” both the room and equipment
load. The storage factors for estimating the load
with the above conditions are contained in
Table 12.
141
CHAPTER 4. SOLAR HEAT GAIN THRU GLASS
SOLAR HEAT - DIRECT AND DIFFUSE
The solar heat on the outer edge of the earth’s
nosphere is about 445 Btu/(hr)(sq ft) on December 21 when the sun is closest to the earth, and
about 415 lStu/(hr)(sq ft) on June 21 when it is
farthest away. The amount of solar heat outside the
earth’s atmosphere varies between these limits
throughout the year.
The solar heat reaching the earth’s surface is
reduced considerably belsw these figures because
a large part of it is scattered, reflected back out into
space, and absorbed by -the atmosphere. The scattered radiation is termed &/fuse or sky radiation,
and is more or less evenly distributed over the
earth’s surface because it is nothing more than a
reflection from dust particles, water vapor and
ozone in the atmosphere. The solar heat that comes
directly through the atmosphere is termed direct
~ndirrtion. The relationship between the total and
the direct and diffuse radiation at any point on
-rth is dependent on the following two factors:
_ 1. The distance traveled through the atmosphere
to reach the point on the earth.
2. The amount of haze in the air.
As the distance traveled or the amount of haze
increases, the diffuse radiation component increases
but the direct component decreases. As either or
both of these factors increase, the overall effect is
to reduce the total quantity of heat reaching the
earth’s surface.
ORDINARY
dow. The direct radiation component results in
a heat gain to the conditioned space only when the
window is in the direct rays of the sun, whereas the
diffuse radiation component results in a heat gain,
even when the window is not facing the sun.
Ordinary glass absorbs a small portion of the
solar heat (5 % to 6y0) and reflects or transmits the
rest. The amount reHected
or transmitted depend?
on the angle of incidence. (The angle of incidence
is the angle between the perpendicular to the window surface anti the sun’s rays, Fig. 18, page 55.)
At low angles of incidence, about 86% or 87oj, is
transmitted and 870 or 9% is reHected, as shown in
Fig. 12. As the angle of incidence increases, more
solar heat’ is reflected and less is transmitted, as
shown in Fig. 13. The total solar heat gain to the
conditioned space consists of the transmitted heat
plus about 40% of the heat that is absorbed in the
glass.
GLASS
Ordinary glass is specified as crystal glass of single
thickness and single or double strength. The solar
heat gain through ordinary glass depends on its
location on the earth’s surface (latitude), time of
day, time of year, and facing direction of the win-
FIG. 12 - REACTION ON SOLAR HEAT (R) , ORDINARY
G LASS, 30” ANGLE OF INCIDENCE
142
L
I’/\RT
1 . LO.\11
U’l‘l
hI,\~I’IN(;
i
s
tyl)ic:tl I’or wootl s;lsli \vintlows. For
met;11 s;141 windows, the glass arca is assumed
CCI~I;I~ to IOO(;{, ol’ t h e
sxli ;~rca Ixxauscthe
c~ontluctivity
of t h e nlctal s a s h i s very high
211d tlic s o l a r heat ai~sorbctl
i n the sash i s
transinittctl almost instant;~neously.
l‘liis
.40x.O6R
Heat &in to Space
= (.40 x .OG R) + .42 R
= ,444 R or .44 R
!
ABSORBED
T
REFLECTED
T
%%!A
Gkf2:
7
.42R
TRANSMITTED
I
Frc. 13 - RUCTION
ON
NOTE: The sash arcn equals
apl~rositnatcly
85:{, o f t h e
masonry opening (or frame
opening with frame walls)
w i t h wootl s a s h w i n t l o w s .
Soft, o f m a s o n r y o p e n i n g
with tl011l)le hung metal sash
w i n t l o w s , nntl 100% of masonry opening
with casement
wintlows.
SOL~K HEA.~ (R) , ORDINARY
G L A S S , 80” A N G L E
OF
INCIDI<NCI:
2 . S o hale iti tlie air.
KOTE:
of the al)sorl~ctl solar llcat ,going i n t o the
space is tlcrivctl from tllc following rca\oning:
The .10”,,
1. The onttloor liim coefficient is approximatelv
2.8 Utu/ (hr)
(sq ft) (deg F) with a 5 mph wintl velocity <luring the
SUlll”VX.
2. The inside film coefficient is approsimately
1.S ntu/ (hr)
(sq ft) (deg F) because, in the average system tlesign, air
velocities across the glass arc approximarely
100-200 fpm.
3. If outtloor temperature
is equal to room temperature. the
glass tcmpcraturc
is alcove I)oth. Thcr-cfol-c al~orl)etl heat
2.8 x 100
.\l)sorl)etl heat flowing out = ___, ,‘i + o,9 = fiO.X~~,. or c;w,.
I
i<
4. .\s tile outtloor tempcratt~re
rises. tile ghs tetllpCl‘;ltlll’C
also rises. tatlsing more of the al~sorl~ctl
heat to flow into
the space. This (an IX ac’c‘ountctl
for 1)~ atitling the trans.
mission of heat across the glass (causctl I)) tcmperatIIre
tlilference I)ctwccn i n s i t l c ant1 nuttlool-s) to the constant
-lOc<, of the al)sorl~ctl heat going insitlc.
5. This reasoning npplics equally well when the outtloor tempcraturc is I~CIOIV the room tcmperatllrc.
Basis of Table 15
- Solar Heot Gain thru Ordinary Glass
Table 25 provides data for 0”, lo”, 20”, SO”, qO”,
and 50”
latitutlcs, for each month of the year and
for each hour of the day. This table includes the
clircct a n d clitfuse radiation and that portion of
the heat absorbed in the glass which gets into the
space. It does not incluclc the transmission of heat
across the glass causctl by a temperature difference
between the outdoor and inside air. (See Chcrpter 5
for “U” values.)
The data in T(rble li is basccl on the following
conditions: -
3. Sea level elevation.
4. A sea level tlewpoint temperature of 66.8 F
(95 F clb, 55 F wb) which approximately corresponds to 4 centimeters of precipitable water
vapor. Precipitable water vapor is all of the
water vapor in a column of air from *a level
to the outer edge of the atmosphere.
If these conditions do not apply, use the correction factors at the bottom of each page of Table 15.
Use of Table 15
- Sotar Heat Gain thru Ordinary Glass
The hold face values
in Table 15 indicate the
maximum solar bent guin for the month for each
exposure. The bold face values that are boxed indicate the yearly maximums for each exposure.
Table 15 is used to determine the solar heat gain
thru ordinary glass at any time, in any space, zone
or building.
To determine the actual cooling load due to the
soln~ heat gain, refer to Chapter 3, “Heat Storage,
Diversity and Stmtificntion.”
CAUTION - Where Estimoting.Mu/ti-Exposure
or Buildings
Rooms
If a haze factor is used on one exposure to determine the peak room or building load, the diffuse
component listed for the other exposures must be
divided by the haze factor to result in the actual
room or building peak load. This is because the
diffuse component increases with increasing haze, as
explainctl on @Se $1.
.
(:H.\l”l‘~l<
4. SOL/\R
IHk:.\‘I‘
(i.\IN
‘I‘HliU
143
(iI..\SS
Example 2 -
.
(Bottom Table 75)
.l’llc
ccllltlilicur on rvllich Y‘d/~/r
1s
Solar heat gain Scptcml,cr 22
\\‘cs t
SOClll~
2:oo
!)!I
I IO
3:oo
I39
XI
7‘01;11
209
“20
2:oo
88
13;
225
3:oo
] ~>‘-> 104
Novemlwr 21
WCS t
South
2:oo
74
I39
3:oo
100
I o-1
Total
21.3
204
4:oo p.m.
149
44
I!):1
Solar heat gain Octolxr 23
\\‘cst
South
Total
226
4:oo p.m.
117
59
156
Solar heat gain 4:oo I’.“‘.
91
59
130
The peak solar heat gain to this room occurs at 3:00 p.m.
on Octolxr 23. The peak room cooling load tlocs not necessarily occur at the same time as the peak solar heat gain,
-
Solution:
By inspection of Table 15 the I~oxed boldface valt:es for
peak solar heat g a i n , occurring at 4:00 p.m. on July 23
= lG4 I!du/(hr)
ci
.
.
/5 i s Islwtl (11, not npl)ly to ’ .
I:intl:
I’cak solar l1cat pill.
Soltltion:
I:rom T,l/J/f,
?
1
*
Solar Gain Correction Factors
(sq fr)
;\ssumc a somewhat Ilazy condition.
.\ltitutlc correction = l.OOi (Iwttom Tnblr IS)
Dewpoint tlilfcrencc = 09.8 - 66.8 = 3 F
Dcwpoint correction = 1 - (3/10 X .Oi) = -979
(bottom Table 15)’
Haze correction = I - .I0 = .90 (Iwttom Table 15)
Steel sash correction = I /.85 (hottom Tnble 15)
Solar heat gain at 4:00 p.m., July 23
= 164 X 1.007 X ,979 X .90 X I/.%
= 171 Btu/(hr)(sq ft)
,’
l-44
I’.\R’I‘ I. 1.0 \I) 1-~5’l’l.\I.\‘i‘l~(;
TABLE 15-SOLAR
O0
0”
HEAT GAIN THRU ORDINARY GLASS
O0
Btu/(hr) (sq ft sash area)
NORTH
Time of Y e a r
J U N E 21
LATITUDE
Exposure
1
AM
6
7
North
Northeast
0
0
East
0
Southeast
South
Southwest
west
I Northwest
0
0
101
101
101
SUN
8
45
II9
II6
9
I
37
b
b
5
5
-
4; I ;; I ii I
MAR 22
FEB 20
J A N 21
Southwest
West
Northwest
lol
101
lo1
0
0
;j
13
13
147
II
II
87
52
II
II
II
II
91
I2
118
167
IIB
12
I2
I2
t
)I I2
100
I2
79
Iii
141
28
I2
t
I2
t
t
12
97
t
-
I8
14
I4
14
14
195
3
4
5
78 1
141
74 t
131
65 t
45 / 0
III
51 0
I3
IO1
I51
IO1
I 13
I3
I3
I3
I63
I3
AC;
Iii
I33
31
I I3
I 13
I 13
1 I50
I4
14
I4
14
lb
223
14
14
I4
I4
43
233
Ii I 1’4 I 1: I I! I
ii
36
52
46
Id Cl ii
411 1 QA 139 152. I21
150
153
118
86
Iii
223
195 1 I51 1 91 1 29
Steel Sash, or
No Sash
X l/.85 o r I.17
Exposure
I4
I4
I4
I4
I4
I3
I2
6
58
31
I4
14
I4
I3
I2
6
107
47
I4
14
I4
I3
12
5
68
31
14
I4
I4
13
I2
6
I 141 141 14 I 14 I 14 ) 13 ) 12 I 6
14
I4
I4
31
68 IO1
IIB
95
14
I4
I4
47
107, 151 1 6 7 1 3 4
14
1 4 1 4 31
68
101
II8
95
210 2 4 0 2 5 0 240 210 I63
100
32
I4
I4
I4
I4
I4
13
12
6
75
is
14
14
14
13
-12
6
;4
ii
I4
13
I2
5
IOJ
;b
102
51
24
I4
I4
13
I2
6
33
34
34
34
33
31
2%
I7
I I41 i4 I 24 1 61 ( 102 1 133 I 141 I 110
1 141 141
14 i 46 ) 103 ) 148-l 163 i 129
1 141 141 14 I 15 / 35 I k-5 I 79 I 67
(2061 2341 245 1234 1206 1 I50 1 97 1 31
141
I IAI
96
124
55
14)
141
141
14 i
14 I
14 1 13 1
I
I? I. 141
43
85
bb
16)
141
II 1
i4 I 14 I 14 I 13 I
II
1 4 I 14 1 1 4 1 1 3 1
II
I
Haze
Altitude
- 15% ( M a x . )
+0.7% per 1000 Ft
Bold Face Values - Monthly Maximums
j
1
East
0 I
(
1
1
1
1
0
0
0
0
0
0
I
.
South
I
FEB 20
&
OCT 23
I
South
Southeast
Eari
Northeast
MAR 22
I
) North
Northwest
west
Southwest
Horizontal
South
southeast
28
I
0
East
APR 20
Northeast
&
North
I
Northwest
1 West
I Southwest
1
Horizonta!
AUG 24
I
I
I
Northeast
JULY 23
I
I
I
Northeast
J U N E 21
North
Northwest
west
Boxed Vaiues - Yearly maximums
I
Horizontal
Dewpoint
Increase From 67 F
- 7 % p e r IO F
M A Y 21
&
North
1 Northwest
I
West
I Southwest
(
Horizontal
37 I 0 I Southwest
Dewpoint
Decrease From 67 F
+ 7% per IO F
’ &
SEPT 22
6 l 0 1 South
I b I o I Southeast
0
East
0
0
0
0
&
N O V 21
6
6
45
II9
II6
j
J A N 21
0
southeast
0
Eest
0
Northeast
0
North
0
Northwest
0
west
1 0 I southwest
0
Horizontal
0
0
0
0
( o
0
0
0
0
0
0
0
0
0
I o
1 0
I o
1 0
DEC 22
South
Southeast
5 1 0 1
south
I 6 I o I Southeast
6
0
East
II
II
6
54
37
1 I53 1 I I8
1 152 I 121
52 1 45
( 91 1 29
I ,j I 7; I I; I 141 14 I 14 I 14 I 13 I II
147
I ,;
4;
4i
I4
I4
I4
I3
II
I56 I54 I33
95
53
20
I4
13
II
b5
74
78 ' 80
82
80
78
74
65
II
I3
I4
20
53
95
133 154 156
II
I3
14
I4
14 I 43
93
I35
147
II I 13 I 141 141 14 I 14 I I5 I 27 I 42 I
87 147 I91
217 2 2 6 217 I91
147
87
1
South
Southeast
East
Northeast
North
Northwest
west
Southwest
Horizontal
Time o f Year
Northeast
1 : 1 North
iI
Northwest
0
west
0
Southwest
1 0 1 Horizontal
34
34
33 I 31 I 28 I 17 I
24
14
14
I3
12
6
14
14
I4
I3
12
6
I4
I4
I4
13
12
6
I4
I4
14
13
I2
5
14
I5
35
65
79
67
I4
46 103 148 lb3 129
24 1 51 I102 II33 I I41 1 II0
245 234 206 I50
97
3I
I4
I4
I4
I3
43
I6
I4
I3
67
56
65
51
‘?3 ) 86 ) 124 1 150
m II I 1 3 I 1 4 1 1 4 1
1’4 I 43 I 96 I I39
I II I I3 I I41 I41 I4 I I4 I I8 I 35 I
91 ( l5i II95 ( 2231 233 1223 ( I95 1 I51
II
6
i4l
33
34
102
61
103
46
35
15
I4
I4
I4
14
14
I4
I I41 141
206 234
II I
13 I
;j I ?C,
I 15;
1;;
153
150
54
61
I I ) I3 1
>I
0” SOUTH LATITUDE
PM
2
14
14
14
191
36
13
13
I3
I3
I51
--I
Solar Gain
Correction
I
14 1 14 1 14 1 13 1 II 1
6 1 0
;i
ii
14
14
14
43
20
53
95 ‘133
I54 -I&. 119
0
217 226 217
191
147
87
28
0
61
55
bb
67
55
55
bl
54.
37
0
I50
124 I ihl
43 I lb I 14 I 13 I II I 5 I 0
9 5 1 iii
14 1 I3 I I I I
5 10
I39-- .-.
I ; I 14 I
;;
31
I;3
148
55
12
I!
I2
I3
12 I I3
II2 1 I3
97
I50
Southwest
west
TIME
Noon
J.g
78
APR 20
II
65 1 74 1 78 t 801 82 1 80 1
156115411~11951 201
I:;
1521
M A Y 21
IO
I
South Lat.
Dec. or Jan.
+ 7%
CM;\1”1‘1:1< 1
.
l-45
SOl..\I< lrI<,\.I‘ C.,\IN ‘l‘H1111CI,.\SS
TABLE 15-SOLAR HEAT GAIN THRU ORDINARY GLASS (Contd)
10°
Btu/(hr)
IO” N O R T H
Time of Year
LATITUDE
,--
Exposure
6
North
Northeast
East
southeast
1 J U N E 21
South
Southwest
west
Northwest
Horizontal
North
Northeast
JULY 23
East
-So”t~~east I.
&
South
_ Southwest
MAY 21
west
Northwest
Horizontal
North
Northeast
AUG 24
East
Southeast
&
South
Southwest
APR 20
west
’ Northwest
Horizontal
SEPT 22
&
MAR 22
OCT 23
&
FEB 20
N O V 21
&
J A N 21
DEC 22
Solar Gain
Correction
19
55
54
18
2
2
2
2
4
5
42
50
26
I
I
I
I
3
I
17
25
18
I
I
I
I
2
AM
i
7
i.8,
44
131
I34
49
8
8
8
8
44
34
127
135
57
7
7
7
7
42
15
II3
138
79
7
7
7
7
38
S
U
N
IO
I
I
Noon
44
106
98
25
I4
I4
I4
14
205
33
109
98
32
14
14
I4
14
210
I5
80
104
60
I4
I4
I4
14
213
43
41
43
44
65
28
14
14
41
14
14
14
14
I4
14
I4
I4
14
14
I4
14
14
14
25
I4
14
41
98
I8
28
65
106
233
243
233
205
31
30
3
I
33
56
22
14
14
43
14
I4
14
I4
I4
14
14
14
14
14
14
I4
I4
14
32
I4
I4
43
98
14
22
56
109
236 247 236 210
14
14
14
15
34
I4
14
14
46
I4
I4
I4
27
I4
14
14
14
14
I4
14
I4
I4
27
60
I4
I4
46
80
1414 3
4
242
250
242
213
‘\
9
50
45
153
140
155
139
55
43
II
13
8
13
8_
13
8
I3
107
166
39
35
148
133
158
142
66
56
I
I 13
II
I3
II
I3
I/
13
107
lb6
16
15
130
III
163
149
94
85
II
13
II
13
II
13
II
I3
I05
I67
I
6
I I I3 14
I
89
103
80
45
. I 130
I64 I51 106
I
97
127
I22
94
I
b
13
I9
24
I
6
II
13
14
I
6
II
13
14
I
6
II
13
14
II 31 97 lb0
1 207
North
Northeast
East
Southeast
South
Southwest
west
’ Northwest
Horizontal
North
Northeast
East
Southeast
South
Southwest
\ West
‘\ Northwest
Horizontal
North
0
5
0
58
0 II8
0 103
0
18
0
5
01
5
0
5
0
22
Northeast
East
Southeast
South
Southwest I
west
’ Northwest
Horizontal
North
Northeast
East
Southeast
South
._ S o u t h w e s t
, west
’ Northwest
Horizontal
Steel Sash. or
No Sash
x l/.85 o r I.17
0
4
9
00
0
0
0,
0
0
0
0
0
0
01
0
0
0
0
0
99
27
99
35
4
4
4
I7
4
I5
86
143
37
I53
65
9
9
9
62
9
28
137
154
74
9
9
9
66
‘5;
4
4
4
I4
12
132
17
161
91
12,
12
12
131
12
I7
130
1631
94
I2
12
12
120
(Max.]
I
I
M
2
3
4
5
4 5 -5&
13
II
13
II
13
II
13
II
43
55
139 155
140 1 5 3
166
107
35
39
13
Jl
I3
II
I3
II
13
II
56
66
142 158
133 148
Ibb 107
I5
16
I3
II
I3
II
I3
II
13
II
85
94
149 l b 3
III
15
130
I67
I05
I2
12
9
12
9
91
65
Ibl
I53
132 143
I7
37
I31
62
12
9
12
9
I2
9
I2
9
94
74
1 6 3 I54
130
137
17
28
120
66
Altitude
+0.7%
Bold Face Values - Monthly
per IO00 F t
Maximums
I2
9
Values
6
SOUTH
;
2
2
18
54
55
4
34
5
7
I
7
i
7 I
I
7
I
57
26
135
50
127
42
42
3
I5
I
7
i
7 I
1
7
7
/
79
I8
I38
25
II3
I7
38
2
6 1
6
6
6
6
97
130
89
3 I ,
5
5
5
5
I8
103
II8 1
58
22
4
1
I
I
I
I
0
0
0
0
0
0
0
0
0
0
South
Southeast
East
Northeast
North
Northwest
‘Nest
Southwest
Horizontal
South
Southeast
East
Northeast
North
.Northwest
west
Southwest
Horizontal
South
4
4
35
99
99
27
I7
4
4
4
4
50
99
86
I5
I4
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
East
Southeast
No&east
North
Northwest
west
Southwest
Horizontal
south
Southeast
East
Northeast
North
Northwest
west
Southwest
Horizontal
- Yearly
LATITUDE
Exposure
South
Southeast
East
Northeast
North
Northwest
west
Southwest
Horizontal
South
Southeast
East
N ortheast
North \
Northwest
west
Southwest
Horizontal
South
Southeast
East
N ortheast
North
Northwest
west
Southwest
Horizontal
44
8
8
8
8
49 I
I34
131
44
Dewpoint
Decrease From 67 F
+ 7% p e r I O F
Boxed
I O”
PM
I3
93
39
14
14
14
14
14
I3
146
109
70
31
I7
96 104
106 104 96
I7
31
70
109,
146
13
I4
14
39
93
I3
14
14
I4
I3
175 202
210
202
I75
13
I4
14
14
I3
I3
14
14
I4
I3
91
42
I4
14
I3
149 121 79 3 6 ) 2 3
109
II6 120 I I6 109
23
36
79
I21
149
I3
I4
I4
42
91
I3
I4
I4
I4
13
lb7
193
202
193
167
13
14
I4
14
17
14
14
47
‘I4
14
56
21
14
27
28
27
14
21
56
14
14
47
I4
14
I7
235 247 235 I
I4
I4
14
14
14
I4
40
I4
I4
81
46
I8
71
73
71
I8
46
81
I4
I4 401
14
14
14
220
230
220
14
I4
I4
E
13
II
13
II
13
II
13
II
19
13
122
127
I51
I64
80
103
I60
97
13
IO
I3
lo
13
10
I3
IO
55
40
149
147
1 4 5 I55
44
66
139
85
Haze
- 15%
T
14
I4
I4
14
24
94
106
45
207
I4
14
14
I4
65
I23
100
28
193
I3
IO
!3
14
66
44
28
155 I45
100
147
149 I23
40
55
65
IO
I3
I4
IO I3
I 14
IO
I3
14
85
139
193
10°
(sq ft sash area)
2
I
I
I
Dewpoint(
Increase From 67 F
- 7% p e r IO F
maximums
Time of Year
DEC 22
.
J A N 21
&
N O V 21
_
FEB 20
&
OCT 23
I
MAR 22
&
SEPT 22
:
L
,
APR 20
1
&
!
AUG 24
:
M A Y 21
i’
&
JULY 23
JUNE 21
South Lat.
Dec. or Jan.
+ 7%
I
l-4
6 I’AR’I I. l.O.\l) l~.4’l’l,\l,\‘l’lN(;
TABLE 15-SOLAR HEAT GAIN THRU ORDINARY GLASS (Contd)
2o”
20”
NORTH
LATITUDE
Time of Year
M A Y 21
SEPT 22
MAR 22
7
Southeast
1 South
Southwest
west
Northwest
Horizontal
I North
Northeast
I
0
I
west
NOV 21
1
I I
19 I
Northwest
Horizontal
North
Northeast
East
Southeast
01
41
4
:
4
0 1 I8 1
0 1 3 I
(
1
I
~~
91 121
9
9
12
I2
68 I127 1
8 1 II 1
0
0
131
I;
1 North
Northeast
1
0)
51 4 8
I
0 I
IO1
7 I II
2
I
I
I Southeast
South
Southwest
0
0
0
0
Northwest
Horizontal
‘74 i;i
7
II
20
7
Ii
I2
7
II
I2
36
92
I35
2
2
2
4
Steel Sash, or
No Sash
Haze
l/.85 o r 1.17
I
-15%
(Max.]
I
Bold
I
2
17 I
19
3
I
Values
28
62
66
73
143
160
I48
122
144 154
176
I21
60
17
23
28
13
I2
8
13
12
8
13
I2
8
13
12
780
79
85
145
lb3
148
ill
138
132
175 II8
55
13
II
IO
13
II
7
13
II
7
I3
/I
7
I4
II
7
I08
II3
89
149 165 142
89
118
III
167
107
48
I3
II
6
I3
II
6
13
II
6
6
13
II
1 38 1 22 I 8 I
140 1 136 1 99 ]
I:49 1 lb3 1130 1
Exposure
I4
I3
13 72 .9
I2
9
-4
4
Time of Year
South
_
28
81
81
II
20
3
3
3
Northwest
west
Southwest
Horizontal
South
Southeast
East
Northeast
North
Northwest
J A N 21
&
3:
75
west
71
Southwest
8
Horizontal
6
South
2
Southeast
2
East
2
Northeast
2
North
29
Northwest
53
west
45
Southwed
;
Horizontal
0
South
0
Southeast
0
East
0
Northeast
0 I North
0 1 Northwest
0 1 West
0
-
_
NOV 21
.
FEB 20
&
OCT 23
MAR 22
.
t
&
SEPT 22
.
South
Southeast
0
1
0
*#I,2 .I‘
100
I3
I71
I41 29
127
52
~~
b8
44
I8
0
Horizontal
Ii
I3
II
8
3
0
South
91
141
91
I3
13
180
I3
i;
46
136
135
4.3
13
172
I7
_
II II
II
100
lb4
127
14
IO1
II
II
II
11
III
lb7
I21
I2
92
88
8
69
144
I28
26
48
7
7
7
7
74
139
II8
18
36
3 3
3
28
73
71
24
5
2
2
2
2
25
59
56
14
4
00
0
0
0
0
0
0
0
0
0
0
0
0
0
n
6
East
Southeast
Northeast
North
Northwest
west
Southwest
Horizontal
_
I3
lb
123
158
91
13
I46
12
12
I3
I
I 1; I i; I 1;
135
I36
46
‘2
6
41 I
20” SOUTH LATITUDE
_
I4
ii
60
I3
I3
I61
97
I,
I2
I3
170
11*.1
I
;;
I46
134
:;
132
159
85
12
135
76
7;
I61
&
J U L Y 23
East
Northeast
North
Northwest
J U N E 21
west
-tauthwrct
__..... __.
Ho+n,+zl
I
+ 7% p e r I O F
I
Maximums
M A Y 21
South
Southeast
Dewrxint
I
Dew&
1 Decrease’From
67 F I Increase ko”,’
+0.7yo p e r IO00 F t 1
- Monthly
.
~~-
5
74 1 II9 1 149 1 lb0 1 I46 1 91
I
Face
I41
4
25 I 33 I
I
APR 20
&
._
0
West
~-~
131
I46
172
I2 I 131
0
0
East
,
;i
1 -3 1 77 1144 IIhA IlGR
Horizontal
X
I
I
I3
13
14
I4
14
I4
49I4
171 1 1961 208 1196
,
Solar Gain
Correction
15
13 1 271
J A N 21
22
Noon
171
I 4 I 9 I 12 I I3
44
52
29
Ii I 141
IOC
99
147 I l41- I
0
Southeast
South
Southwest
FEB 20
IO
251
I
0
East
&
9
PM
T I M E
4
I;
14
14
14
ii
1;
44
9
12
14
i4
14
I4
41
96
9
12
14
14
14
15
38
83
60
121
176 216 232 250 232 216
28
23
17
15
I4
14
I4
I5
132 138
III
73
31
14
I4
I4
I48
lb3
145
99
46
I4
14
14
70
85
79
57
29
I4
14
I4
8
I2
13
I4
I4
14
I4
14
8
12
I3
14
14
14
29
57
8
12
13
14
14
14
46
99
8
I2
I3
I4
1 4 14 31
73
8
55
II8
175 216
240 251 240 216
6
IO
II
I3
I4
14
14
14
I4
45
111 118 89
50
18
14
14
14
53
142
lb5
149 106
51
I4
14
29
89
113 I08
98
55
lo”
I4
14
2
7
II
14
20
24
26
24
20
2
7
II
I3
14
14
20
55
98
2
7
II
13
I4
14
I4
51
I06
2
7
II
I3
I4
14
14
18
50
5
48
107 167 210 235 247 235 210
0
6
II
I3
14
14
I4
14
14
0
83
87
59
22
I4
I4
I4
14
0
I30
lb3
149
104
45
I4
14
14
0
99
136 140
120
84
41
I5
I4
1 0 1 8 1 22 1 38 t 52 1 63 t 65 1 63 I 52
1 0 1 6 t I I t I3 1 I41
I51 41 1 84 1 120 t
1 0 1 6 1 I I I 13 1 14 1 I41
I4 1 45 II04 1
East
O C T 23
8
3
3
3
II
20
71
75
31
3
3
3
3
Southeast
South
Southwest
west
Northwest
Horizontal
North
Northeast
APR 20
6
S U N
Southwest
west
Northwest
Horizontal
East
&
AM
28 41
33
81
154
144
81
148
IbO-
North
Northeast
A U G 24
1
North
Northeast
East
North
Northeast
East
Southeast
South
Southwest
west
Northwest
Horizontal
&
&
I
Exposure
J U L Y 23
DEC
2o”
Btu/(hr) (sq ft sash area)
Boxed
1 I
Values
- Yearly
7%
maximums
per
67 F
Iii’
I
I
South Lat.
n-r w J a n .
--+:‘70
(:tl,\I”I‘ER ,k.
SOl>.\l<
ill:.\~I’ <;/\lN ‘I‘[-llilJ
TABLE 15-SOLAR
HEAT GAIN THRU ORDINARY GLASS (Contd)
3o”
30”
3o”
Btu/(hr) (sq ft sash area)
NORTH
T i m e o f Year
J U N E 21
J U L Y 23
&
M A Y 21
A U G 24
&
A P R 20
SEPT 22
&
M A R 22
O C T 23
&
FEE
147
(;l,,\SS
20
N O V 21
&
J A N 21
DEC 22
Solar Gain
Correction
LATITUDE
A M
S U N T I M E
b
7
8
9
North
Northeast
East
Southeast
South
Southwest
west
Northwest
Horizontal
North
Northeast
I East
Southeast
South
Southwest
west
Northwest
Horizontal
North
Northeast
East
Southeast
South
Southwest
west
Northwest
Horizontal
North
Northeast
East
Southeast
South
Southwest
west
Northwest
Horizontal
33
105
108
42
5
5
5
5
I9
22
93
100
42
4
4
4
4
I5
6
55
66
37
2
2
2
2
b
0
~ 0
0
0
0
0
0
0
0
29
139
156
75
IO
IO
IO
IO
61
20
131
I55
82
9
9
9
9
bb
8
108
147
98
8
8
8
8
47
5
74
124
98
9
5
5
5
25
18
130
lb1
90
12
I2
12
12
131
I4
123
lb4
100
I2
I2
I2
12
123
II
100
lb5
127
I3
II
II
II
107
IO
90
158
131
18
IO
IO
IO
81
14
97
143
90
I4
14
14
I4
180
13
89
I45
100
I4
I3
I3
I3
I76
I3
66
148
129
27
I3
13
I3
Ibl
I2
40
I.44
152
60
I2
12
12
I35
North
Northeast
East
Southeast
South
Southwest
west
Northwest
Horizontal
North
Northeast
East
Southeast
South
Southwest
west
Northwest
Horizontal
North
1 Northeast
East
Southeast
South
.O
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
3
33
79
73
18
3
3
3
6
I
8
27
28
IO
I
I
I
2
0
Exposure
1
01
1
Steel Sash, or
No Sash
X Ii.85 o r I.17
Bold
01
0 1
0
0
IO
Noon
I
2
14
14
14
17
21
17
14
14
250
14
14
I4
22
30
14
14
I4
246
I4
I4
14
39
63
39
14
I4
135
14
14
I4
b7
105
67
14
14
212
14
14
14
14
19
44
44
19
240
I4
I4
14
14
27
53
44
16
236
14
14
I4
I5
58
82
46
I4
225
14
14
I4
25
98
II3
48
14
202
14
14
14
14
15
73
98
55
217
14
I4
I4
14
20
83
99
46
214
13
13
13
13
47
II2
102
27
200
I3
I3
13
13
82
I41
LO3
15
179
8
II
I2
13
I4
39
I8
I2
I3
I4
135
132
94
43
14
142
lb3
I59
136
92
57
92
I21
139 145
8
II
IS
47
92
8
II
I2
I3
14
8
II
I2
13
14
49
100
143
171
179
b
9
II
I2
I2
lb
9
II
I2
I2
109 II6
83
35
I2
127 lb1
lb2
143
104
68
109 137
154
159
6
9
23
64
104
b
9
II
I2
I2
6
9
II
I2
I2
27
71
109
l3b
145
4
9
II
I2
I2
10)
91
II)
121
12)
0 ( 92
105
80
32
I2
0 114
I57
l b 2 143
108
0
64 II3
142 I59
lb3
9
28
72
108
9
II
I2
I2
13
13
I3
47
139
136
43
I3
171
I2
12
I2
64
I54
143
35
I2
I36
I2
121
I2
72
I59
143
32
Haze
-15%
Face
Altitude
(Max.)
Values
II
14
14
55
19
98
44
73
44
I5
19
I4
I4
I4
14
14
14
217 240
14
14
46
I6
99
44
83
53
20
27
14
14
14
I4
I4
14
214 23b
I3
I4
27
14
102
46
II2
82
47
58
I3
15
13
I4
I3
14
200 225
I3
I4
15
I4
103
48
I41
II3
82
98
I3
25
13
14
13
14
179 202
+0.7% p e r 1 0 0 0 F t
- Monthly
Maximums
30”
PM
3
14
14
14
14
14
90
143
97
180
13
I3
13
I3
I4
100
145
89
I76
I3
13
I3
I3
27
129
I48
66
lbi
I2
I2
I2
12
60
152
144
40
I35
I2
II
12
II
12
II
I5
II
121 92
159
lb3
94 132
12
I8
143 100
II
9
II
9
I I
9
23
9
137
109
lb2
lb1
83 II6
II
9
109
71
II
9
II)
9
II
9
28
9
142
I I3
lb2
I57
80 105
Values
South
Southeast
East
Northeast
North
Northwest
west
Southwest
Horizontal
South
Southeast
East
Northeast
North
Northwest
West
Souihwest
Horizontal
South
Southeast
East
Northeast
North
Northwest
8
8
8
8
57
142
135
39
49
6
b
6
b
68
127
109
lb
271
4
4
3
3
3
3
18
73
79
33
6
I
I
I
I
IO
28
27
8
2
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
South
Southeast
East
Northeast
North
Northwest
west
Southwest
Horizontal
South
Southeast
East
Northeast
North
Northwest
West
Southwest
Horizontal
South
Southeast
19
0
I
0
Horizontal
- Yearly
LATITUDE
Time of Year
EXpOSUre
18
29
33
12
IO
5
I2
IO
5
12
IO
5
I2
IO
5
90
75
42
lb1
I56
108
130139105
I31
61
19
14
20
22
I2
9
4
12
9
4
I2
9
4
12
9
4
100
82
42
lb4
155
100
123
131
93
123
66
I5
II
8
6
II
8
2
II
8
2
II
8
2
13
8
2
127
98
37
lb5
147
6b
100
108
55
IO7
47
b
IO
5
0
IO
5
0
IO
5
0
IO
5
0
I8
9
0
I31
98
0
158
124
0
90
74
0
81
25
0
Dewpoint
Decrease From 67 F
+ 7% p e r IO F
Boxed
6
4.5
SOUTH
DEC 22
J A N 21
&
N O V 21
FE6 2 0
&
OCT 23
w e s t
Southwest
Horizontal
South
Southeast
East
Northeast
North
Northwest
west
Southwest
Horizontal
Dewpoint
Increase From 67 F
- 7 % p e r IO F
maximums
MAR 22
&
SEPT 22
APR 20
&
AUG 24
MAY 21
&
JULY 23
1
South Lat.
Dec. or Jan.
+ 7%
’
l’;\l<-l‘ I, 1.0 \I) I~:4l‘IxI.\‘I 1x(;
l-48
TABLE 15-SOLAR
HEAT GAIN THRU ORDINARY GLASS (Contd)
Btu/(hr) (sq ft sash area)
40”
N O R T H--~I
LATITUDE 1
S U N
AM
Time of Year
J U N E 21
8
/
JULY 23
&
M A Y 21
AUG 24
&
APR 20
North
Northeast
East
Southeast
&uth
Southwest
‘west
Northwest
Horizontal
North
1
32 1
Southeast
,5outh
Southwest
/West
Northwest
Horizontal
North
Northeast
East
Southeast
/South
Southwest
/West
Northwest
Horizontal
I
1
1
IO
II
Noon
I
20 1
51 1
b
b
6 I
61
3l\(
24 1
I
1
1
)
9
T I M E
54 I
51
5
5
5
24
7
68
84
48
3
3
3
3
9
PM
2
3
12
13
14
14
14
14
14
13
112
73
30
I4
14
14
I4
I3
95
44
14
I4
I4
13
162 1 I4
88 I 109 [II:
99
71
34
14
I4
131
IO
12
19
35
44
54
44
35
19
IO
I2
I3
14
I4
34
71
99
III
IO I 12 I I3 I 14 I I41
14 I 44 I 95 I142 I
IO)
121
I31 I41 14j I41 301
82 I134 1179 210 232 23,7T232 210 179
14 1 I2 / I3
14
14
14
14
14
13
6
26
14
14
14
I4
1.7
43
14
14
I3
96 I II9 i-1245
1::
82
1442
I5
I4
I3
101
131 261
441 b3/
691 631
441 261
IO
I2
I3
14
I5
42
82
110 I25
IO
I2
13
I4
I4
14
43
98 144
IO
I2
13
I4
I4
I4
14
26
66
73
I25 171
203 225 233 225 203
I71
8
II
13
I4
I4
I4
I4
I4
I3
102
82
46
I6
14
14
14
I4
13
147 I62
I45
IO1
45
IS
14
14
I3
I05
I38
I46
139 107
bb
25
I4
I3
8
24
51
89
97 102
97
89
51
8
II
I3
I4
25
66
107 139 I46
8
II
13
14
14
I4
45
101
145
8
II
13
14
I4
I4
I4
lb
46 ’
47 100 150 185 205 214 205 I85
150
4
14 0 ” S O U T H L A T I T U D E
5
6
EXpOSW
12
20
32
12
IO
6
12
IO
6
12
IO
6
12 1 IO 1 6
IO 9 1 88 151
I62 1 lbl i I26
73~11211331118
134
82
31
I2
14
24
I2
IO
5
I2
IO
5
12
IO
5
131
101
51
I19
96
54
I64 lb1
118
105
127 106
126
73
24
II
8
7
II
8
3
II
8
3
II
8
3
24
8
3
138 105
48
I62
147
84
82
102
68
100
47
9
Time o f Year
South
Southeast
East
Northeast
North
1
1
1
1
Northwest
west
Southwest
Horizontal
South
Southeast
East
Northeast
North
Northwest
west
Southwest
Horizontal
1
DEC 2 2
J A N 21
&
1
NOV 21
.
South
Southeast
East
Northeast
FE6 2 0
North
&
Nokthwest
Wed
Southwest
Horizontal
OCT 23
.
North
Northeast
East
SEPT 22
1
Southeast
/‘South
Southwest
/west
Northwest
Horizontal
&
MAR 22
North
Northeast
East
OCT 23
Southeast
rSouth
&
Southwest
FEE 2 0
1
I
1
, west
Northwest
Horizontal
North
Northeast
East
Southeast
/South
Southwest
/west
N O V 21
&
JAN 21
Northwest
Horizontal
North
Northeast
East
southeast
/South
Southwest
DEC 22
/west
Northwest
Horizontal
Solar Gain
Correction
1
Steel Sash, or
No Sash
X li.85or I.17
1
I
1
O(Il6ll491139( 991 451
0
I57
133
0
II0 122
0
5
9
I2
I4
41’
0
5
9
I2
I3
13
0
5
9
12
13
I3
0
21
67
124 I53
I76
I41
90
140
90
14
14
183
131 131 121
91
41
I4
12’
9
122 I10
81
44
I33
157 162 144
45
99. 139 149
I3
13
26
58
I76
I53
124
67
0
2
IO
II
I2
I2
I2
II
IO
6
0
35
3:
I2
II
12
I2
12
II
IO
b
01 851 11711221 881 391
121 I21
I I 1 IO1
61
0 i 81 1 I32 I Ibl I I63 I 1441
107 I 6 3 I 20 1 101
61
01 211 5911041 1371 1541
I62 II541 1371 1041 591
0
2
6
IO
20
63
107 144 I63
lb #I / 1321
0
2
b
IO
II
12
I2
39
88
12 211171
0
2
b
IO
II
I2
I2
12
II
I2
33
0
8
29
64
101
123 129
123 I01
64
29
0
0
3
7
9
IO
II
IO
9
7
3
0
0
I2
7
9
IO
II
IO
9
7
3
0
0
91
100
74
33
II
IO
9
7
3
0
0 109 144 I56 144 II6
70
27
7
3
0
0
59
104 I39
158 I66
I58
139 104
59
0
0
3
7
27
70 II6 144 I56 144 109
0
0
3
7
9
IO
II
33
74 100
91
0
0
3
7
9
IO
II
IO
9
7
I2
0
0
I6
43
73
92
I03
92
73
43
I6
0
0
2
6
9
IO
IO
IO
9
b
2
0
0
7
6
9
IO
IO
IO
9
b
2
01
01 72 1 861 681
311
101 101
91
61
21
0 I
0 1 88 I 1341 I48 I 1421 I I5 1 7 3 1 3 0 t 7 1 2 1
01 01 51 1 991 I341 I581 I65 I I581 1341 991 51
0
0
2
7
30
73
II5 142 148
I34
88
0
0
2
6
9
IO
IO
31.
68
86
72
0
0
2
b
9
IO
IO
IO
9
6
7
0
0
8
32
55
76
85
76
55
32
8
Haze
-15% ( M a x . )
Altitude
+0.7% per 1000 Ft
Bold Face Values L Monthly Maximums
Dewpoint
51
5
I2
95
II6
51
21
01
East
0
0
0
0
0
0
Northeast
North
Northwest
1
Decrease From 67 F
+7%per IOF
01
0
0
0
0
&
SEPT 22
Southwest
Horizontal
South
Southeast
East
Northeast
North
Northwest
West
/
APR 20
1
%
A U G
2 4
M A Y 21
East
Northeast
North
&
west
JULY 23
Northwest
Southwest
Horizontal
South
Southeast
East
0 1 Northeast
0
0
0
E
l
west
2
0
2
0
2 1 0 1
2 I
0 I
21 I 0 1
81 1 0 1
851 01
35
8
0
0
0
0 I
0
0
:
0
0
0
0
0
0
0
0
0
i
0
0 I 0 I
0 1
MAR 22
North
Northwest
west
I
J U N E 21
Southwest
Horizontal
Dewpoint
Increase From 67 F
- 7 % p e r IO F
Boxed Values - Yearly maximums
South Lat,
Dec. or Jan.
+ 7%
TABLE 15-SOLAR HEAT GAIN THRU ORDINARY GLASS (Contd)
50°
50”
50°
Btu/(hr) (sq ft sash area)
NORTH
LATITUDE
Time of Year
JUNE 21
Exposure
1
1
AM
6
North
Northeast
East
Sod east
South
Southwest
west
Northwest
Horizontal
29
126
I39
64
a
8
8
a
44
i
JULY 23
&
M A Y 21
-AUG 24
&
APR 20
SEPT 22
&
MAR 22
OCT 23
&
FEB 20
I-JOV 21
&
J A N 21
S U N
101
I I
14
lb
94
124
b8
14
I4
14
197
14
14
41
98
87
23
I4
I4
214
North
Northeast
East
Southeast
South
Southwest
west
Northwest
Horizontal
21
I I4
I31
65
I4
I5
961
136
6
80
North
Northeast
East
Southeast
South
Southwest
west
Northwest
Horizontal
North
Northeast
East
Southeast
South
Southwest
west
Northwest
Horizontal
a
76
T I M E
1Noon
14
14
431
109
98
14
14
I4
bl
93
61
I4
I4
220
14
14
I41
70
I06
98 1 451
I53 I 1321
4
5
I
I4
14
14
14
68
124
94
lb
197
I3
13
13
I3
39
135
136
50
173
I2
I2
I2
I2
lb
I26
lb2
94
133
12
IO
IO
IO
IO
102
lb4
125
86
I4
I4
141
14
80
I3
I3
I31
I3
50
I2
I2
I21
12
21
6
Exposure
29
8
South
Southeast
East
Northeast
North
Northwest
West
Southwest
a
8
8
64
139
126
44
Ii
IO
101
IO
IO
North
Northeast
East
Southeast
South
Southwest
west
Northwest
Horizontal
C
0
C
C
C
C
C
C
C
North
Northeast
c
c
South
Southeast
East
Northeast
North
or
J A N 21
&
4 1 East
4 1
FEB 20
Northeast
OCT
23
’
I
12
1 121
12
12
1
1 12
1 IO
I 8
I 4
1 0
1 Southeast
MAR 22
&
SEPT 22
]
01
East
0 I
0
0
Northwest
west
Southwest
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
South
Southeast
East
Northeast
North
Northwest
west
Southwest
Horizontal
0
0 l
0
0
0 l
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
144 1157 1145/ill
I
35 1 79 II05 / 9 9
)
6
6
57
127
II6
21
6
6
30
5
5
47
8
8
28
127
143
b7
8
8
47
6
6
23
99
25
5
5
I9
Sash,
,
DEC 22
Horizontal
21
6
61
6
6
14 1 I4 1 13 1 12 1 10 1 8
89 I 40 I 13 I 12 I IO I
8
107 I II61
N o Lash
l/.85 o r 1.17
ime of Year
&
0
0
“0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
X
14
I4
I41
26
98
3
NOV 21
94
53
4
4
4
4
I3
North
Northeast
East
Southeast
South
Southwest
west
Northwest
Horizontal
Steel
-
I4
I4
I4
23
87
98
41
14
214
1 50” SOUTH LATITUDE
PM
2
b
6
b
33
DEC 22
Solar Gain
Correction
I
1
I31
62
b
b
33
9
9
9
107
153
107
9
9
53
7
7
7
8
8
8
67
143
127
28
8
47
6
6
6
( M a x . ) 1 +0.7%
II6
23
6
33
107
47
5
I9
41
27
3
5
I
I
I
I
34
62
51
5
4
0
0
Dewpoint
Decrease From 67 F
per 1000 Ft
Bold Face Values - Monthly Maximums
4
4
4
4
70
95
64
4
I3
3
3
3
100 I 62 I 25 I
3 I
I41
I31
99
31
100
7
7
40
Altitude
-15%
6
6
6
21
II6
127
57
6
30
5
5
5
691
73
29
Boxed
+ 7 % p e r IO F
Values
- Yearly
,
l
&
South
Southeast
East
Northeast
North
Northwest
AUG 24
MAY 21
&
JULY 23
J U N E 21
west
Southwest
Horizontal
Dewpoint
Increase From 67 F
- 7% per 10 F
maximums
APR 20
South Lat.
Dec. or Jan.
+ 7%
Heat C;aitl to Space
= (.4 x 52 R) + .43 R
= .G38 R, or .64 R
REFLECTED
I
.4‘3 R
TRANSMITTED
.OBx.51:.77R
/
/
ITT-t-.4 x.i5x .51 x.77R
.
ALL GLASS TYPES-WITH AND WITHOUT
SHADING DEVICES
Glass, ollw tltnn odi~nm-y
solar heat because it
1. May be thicker, or
glass, absorbs more
2. May be specially treated to absorb solar heat
(heat absorbing glass).
These special glass types reduce the transmitted
solar heat but increase the amount of absorbed
solar heat Nowing into the space, Normally they
reflect slightly less than ordinary glass because part
of the reHection takes place on the inside surface.
/I portion of heat reflected from the inside surface
is absorbed in passing back through the glass. The
overall effect, however, is to r-educe the solar heat
gain to the conditioned space as shown in Fig. 15.
‘Refer t o Ilerr~ S, @ge 51, for absorptivity, reflecivity and transmissibility of common types of glass
at 30” angle of incidence.)
The solar heat gain factor through 52% heat
absorbing glass as compared to ordinary glass is
.64R/.88R = ,728 or .73. This multiplier (.73) is
used with Table 15 to determine the solar heat gain
thru 52(,v0 heat absorbing glass. Multipliers for various types of glass are listed in Table 16.
The effectiveness of a &ndi~g device depends
on its ability to keep solar heat from the conditioned space. .\I1 shading devices reflect and absorb
a ma,jor portion of the solar gain, leaving a small
portion to be tr-ansmitted. The outdoor shading
devices are much more effective than the inside devices because all of the reflected solar heat is kept’
out and the absor!~etl heat is dissipated to the outdoor air. Insirle devices necessarily dissipate their
absorbed heat within the conditioned
space and
.
Heat Gain to Space
= (.40 X .15 R) + (.37 X .77 R) + (.I!2 X .77 R)
+ (.08 x 31 X .77 R) + (.40 X .I5 X .51 X .77 R)
= ,492 R or .49 R
FIG. 16 - REACTION
ON
SOLAR
HEAT (R)
P L A T E GL A S S , W H I T E V E N E T I A N
OF
B L I N D,
, I/~-INCH
30”
IN C I D E N C E
1
AN G L E
must also reflect the solar heat back through the
glass (Fig. 16) wherein some of it is absorbed. (Refer
to Item S, @ge 51, for absorptivity, reflectivity and
transmissibility of common shading devices at 30”
angle of incidence.)
The solar heat gain thru glass with an inside
shading device may be expressed as follows:
Q = [.4ag +
tg (%d f kd f r$sd f .$$,d)] ,$
where:
Q = solar heat gain to space, Btu/(hr)(sq ft)
R = total solar intensity, Bttt/(hr)(sq ft), (from T&k
a = solar absorptivity
t = solar transmissibility
r = solar reflectivity
g = glass
sd = shading device
38 = conversion factor from Fig. I2
IS)
For drapes the above formula changes as follows,
caused by the hot air space bettveen glass and drapes:
Q
= [.?hp + tg (.85&l + tsd + rgrSd + .%,r,d)] -&
The transmission factor U for glass with 100yO
drape is 0.80 Rtu/(hr) (sq Et) (F).
The solar heat gain factor thru the combination
in Fig. fh as compared to ordinary glass is .49R/.88R
=.557 or .56. (Refer to Table 16 Eor r/-inch regular
plate glass with a white venetian blind.)
NOTE: /\ctually the reaction on the solar heat reflcctctl Ijatk
throltgh the glass from the hlintl is not always itlcntical to the lirst pass as assumed in this example. The
first pass through the glass filters ant most of solar
radiation that is to he al)sorhetl in the glass, and the
second pass al)sorbs somewhat less. For simplicity. the
react ion is assumed identical, since the quantities are
normally small on the second pass.
(i. Outdoor canvas awnings ventilated at sides and
top. (See T/lb/e 16 I’ootnote.)
7. Since Tfr/Ile Ii is basctl on the net solar heat
gain thru ordinary glass, all calculated solar
heat factors are clividetl by .88 (Fig. 12).
8. ‘1‘11~
Basis of Table 16
Over-all
Factors for Solar Heot Gain thru Glass,
With and Without Shading Devices
Use of Table 16
-Over-all Factors for Solar Heat Gain thru Glass,
With and Without Shading Devices
The factors in TnOle 16 are multiplied by the
values in Table I5 to determine the solar heat gain
The Eactors in Tcrblc 16 ~lre I,aserl on:
!. ,411 o u t d o o r film coefficient o f 2 . 8 Btu/(hr)
(sq ft) (deg F) at 5 mph wind velocity.
2. An inside film coefficient of 1.8 Btu/(hr)(sq ft)
(deg F), 100-200 fpm. This is not 1.47 as normal!y used, since the present practice in well
designed systems is to sweep the window with
a stream of air.
3. A 30” angle of inciclencc which is the angle
at which most exposures peak. The 30” angle
oE incidence is approximately the balance
point on reduction of solar heat coming
through the atmosphere and the decreased
transmissibility of glass. Above the 30” angle
the transmissibility of glass decreases, and
below the 30” angle the atmosphere absorbs
or reflects more.
4. All shading devices fully drawn, except roller
shades. Experience indicates that roller shades
are seldom fully drawn, so the factors have
been slightly increased.
5. Venetian blind slats horizontal at 45” and
shading screen slats horizontal at 17”.
TYPES OF GLASS OR
SHADING
DEVICES*
Ordinary Glass
Regular-Plate,
IA”
Glass, Heat Absorbing
Venetian Blind, Light Color
Medium Color
Dark Color
Fiberglass Cloth, Off White (5.72 - 61/58)
Cotton Cloth, Beige (6.18 - 91/36)
Fiberglass Cloth, Light Gray
Fiberglass Cloth, Tan (7.55 - 57/29)
Glass Cloth, White, Golden Stripes
i
Fiberglass Cloth, Dark Gray
Dacron Cloth, White (I.8 - 86/81)
Cotton Cloth, Dark Green, Vinyl Coated
‘-(similar to roller shade1
&ton Cloth, Dark Greek (6.06 - 91/36)
I
average absorptivity, reflectivity and trans-
missability for common glass and shading devices at a 30” a. ngle of incidence along with
shading factors appear in the table below.
thru diferent combinations of glass and shading
devices. The correction factors listed under Table 15
are to be used if applicable. Transmission due to
temperature diffcrcnce between the inside and outdoor air must be added to the solar heat gain to
determine total gain thru glass.
Example
3 -
blind
on
inside,
Find:
Peak solar heat gain.
Solution:
By inspection of Table 15, the boxed boldface values for
peak solar heat gain, occurring at 4:00 p.m. on July 23
I
= 164 Btu/(hr)(sq
f:)
Reflectivity
(4
.nx
is
ii
.51
.39
.27
.60
.51
.47
Transmissibility
(9
.86
.77
(1 - .05 - a)
.12
.03
.Ol
.35
.23
.23
.14
.54
.42
.41
29
.0!2
.28
Jvi
.--
texture: figures in parentheses are ounces per sq yd, and yarn
count warp/filling. Consult manufacturers for actual values.
/
Shades
Given:
West exposure, 40” North latitude
Thermopane window with white Venetian
:3/, drawn.
.60
rne actual drapery material may be different in color and
Drawn
Occasionally it is necessary to estimate the cooling load in
a building where the blinds are not to be fully drawn. The
procedure is illustrated in the following example:
.-\bsorptivity
(4
36
.iia
by mfg.
.37
.58
.72
.05
.2G
.30
.44
.05
.02
-‘factors tor various draperies are given for guidance only since
. .
Partially
-:::
14
.-.28
.oo
.70
tcompared to ordinary glass.
$For a shading device in combination
Solar Facto*
1 .oo
.94
- .56;
.651
.75i
.482
.56i
.59+,
.64f
.65,+
.75f
.76$
.88$
.76f
with ordinary glass.
_
1-53
I’,\liT I. I.O:\II ES-l‘lhl.\TING
Thcrnwlx~rw wintlows have n o sasll; t h e r e f o r e , s a s h arca
correcli~ln = 1 /.%I (hottom Tn/r/e 15)
I n
t h i s
vcncti;tn
exnmplc,
7/, o f
the wintlow i s ~ovcrctl with tllc
Ijlintl anti I/ is n o t ; Illcrcforc.
F a c t o r
factor.
eq”“ls ‘V, o f
t h e ‘iol3r hc;lt pin
(IlC 0vc1.311 fxtor -t- I/, of the gl;1ss
F a c t o r f o r :+‘, tlrawn = (‘fi X 3)+ (IA X .80)(Tn/~k 1 6 )
= 59
59
= 164 X _Hr,
Solar heat gain
= II4 l)LU/ (hr) (“‘I ft) .
Example 4 -
Pea& Solar Heat Gain thru Solex “R”
Glass
Given:
Weste x p o s u r e , ,40” N o r t h laritutle
IA”
S&x “R” glass in steel sash, double
hung
wintlow
TABLE 16-OVER-ALL FACTORS FOR SOLAR HEAT GAIN THRU GLASS
WITH AND WITHOUT SHADING DEVICES*
Apply Factors to Table 15
Outdoor
wind
velocity,
TYPE OF GLASS
ORDINARY
GLASS
’ REGULAR PLATE
(i/4 inch]
HEAT ABSQRBING, GLASStt
40 to 48%, Absorbing
f ,-i 48.to 56% ~At&binq
.,
56 to 76‘)” Absorbing
DOUBLE PANE
.
.! Ordinary Glass
Regular Plate
48 to 56% Absorbing outside;
Ordinary Glass inside.
1-48 to 560/d Absorbing outs&:
R.eguler Plate inside.
5
I
mph
Angle
SEE
_’
I .oo
of
incidence,
- Shading
30”
INSIDE ’
VENETIAN BLIND*
CLASS
FACTOR
..-
I
Medium
Light
_ _ _
Dark
C o l o r Cnlor.
--.-. I Calor
.b5
1
8
.I5
.94
.56
.b5
.74
.80
.73
.56
.53
.b2
.72
.I2
.59
.b2
.b2
.51
.54
.5b
.I I
.I0
.90
.80
.54
.52
.bl
.b7
.I4
.59
.b5
.I2
.52
)
.36
.50
.83
.b9
PAINTED GLASS
.-Light color
,’ ‘M&dium Co&r
Dark Color .
.28
.39
.50
STAINED GLASS+*
.70
36
.bO
.32
.4b
.43
appear on next page.
A8
.47
)
.39
)
.39
.43
.%I
.52
.b4
.57
f
.I0
.I2
.I0
vent. sides & top
Light
Color
Color -,I
1
__--
.I3
.22
.I5
.20
.25
.I2
.2 I
.I4
.I9
.24
.I I
.I0
.I2
.I I
.I6
.I5
.I0
.I8
.I6
.I4
.I0
.I2
.20
.I8
.I6
.I2
.I I
.20
.I8
.I4
.12
.I8
.Ib
.22
.20
.I0
.I3
.I0
I
window
OUTSIDE
1
A W N I N G S
17O horiz. slats
Inside
.43’)
I
covering
Light on
Outside Medium’‘* Dark 8‘
Dark on
Color
Color
Liaht
I Color
.75 I
fully
OUTSIDE
SHADING
SCREENt
45’ horir. slats
or ROLLER ‘SHADE
.:_ ,56 : ;
devices
OUTSIDE
VENETIAN BLIND
45’ horiz. or vertical
.I4
-TRIPLE P A N E
Ordinary G&s
Regulbr Plate
Footnotes for Table 16
-
1
.I0
.I!
.I0
.I !O
.I I
.I0
.I0
.I2
.I I
.I0
.I8
.I5
.I2
.I0
.I6
.I4
.20
.I7
CM,\l’TI:I<
4.
SOL,\li f-I1:,\7‘
C,,\IN
1-53
‘1‘111~~1 C;I>;\SS
Pig. 15 md Ih, (2) by applying the absorptivity, rellectivity and transmissibility of glass and shades
listed in the table on Page 51, or determined Crom
manul’acturcr,
and (3) by distributing heat absorbed
within the dead air space and glass panes (rig. 17).
Example 5 - Approximation of Over-d Factor
Given:
,A cornl)inatinn as in Fig. fh backed on the inside with another pane of IA-inch regular plate glass.
Find:
The
G.
I’i - REACTION
ON
S OLAR HEAT
(R) , ~-INCH
P LATE GLASS, WHITE V ENETIAN BLIND , ~/-INCH
P LATE GLASS,
30” A NGLE
OF
INCIDENCE
APPROXIMATION OF FACTORS FOR COMBINATIONS
NOT FOUND IN TABLE 16
Occasionally combinations of shading
types of glass may be encountered that
ered in Table 16. These factors can be
(1) by using the solar heat gain flow
devices and
are not covapproximated
diagrams in
over-all
factor.
Solution:
FiczLr’e 17 shows the distribution of solar heat. The heat al).
sorbed I,etween the glass panes (dead air space) is divided
45% and 557” respectively between the in and out flow. The
heat ahsorbed within the glass panes is divided 207” in and
80%) out for the outer pane. and 7.57” in and 25% out for
the inner pane. These divisions are based on reasoning partially stated in the notes under
13, which assume the
outdoor film coefficient of 2.8 Btu/ (hr) (sq ft) (deg F), the
indoor tilm coefficient of 1.8 Rtu/ (hr) (sq ft) (deg F). and the
over-all thermal conductance of the air space of 1.37 Btu/
(hr) (sq ft) (deg F).
Heat gain to space (Fig. 17)
= (.75 X .I5 X .I2 X .77R) + (.77 X .I2 X .77R)
+ .45 [(.37 X .77R) + (.08 X .5l X .77R)
+ (.08 X .I2 X .77R)]
+ .20 [(.l5R) + (.I5 x .5l X .77R)]
= .2G84R or .27R
gain factor as compared to ordinary glass
= .27R/.88R = .3l
Solar heat
Equotionr: Solar Gain Wifhouf Shades = (Solar Data from Table 15) X (Glass Factor from table)
Solar Gain With Shader = (Solar Data from Table 15) X (Over-all Factor from table)
Solar Gain Wifh Shades Partially Drown = (Solar Data from Table 15) X
[(Fraction Drawn X Over-all Factor) + (1 - Fraction Drawn) X (Glass Factor)]
Footnotes for Table 16:
ding devices fully drawn except roller shades. For fully drawn
roller shades, multiply light colors by .73, medium colors by .95, and
dark colors by 1.08.
tFactors for solar altitude angles of 40’ or greater. At solar altitudes
below 40”, some direct solar rays pars thru the slats. Use following
multipliewMULTIPLIERS FOR SOLAR ALTITUDES BELOW
**Commerdol
per inch.
ttMost heat absorbing glass used in comfort air conditioning is in the
40% to 56yo range; industrial applications normally use 56yo to
7Oyo. The following table presents the absorption qualities of the
most common glass types:SOLAR
40’
Trade
Name or
Description
6:00 a.m.
6:00
p.r”.
6~45
5:15 a.m.
p.m.
7:30
4:30 a.m.
p.m.
5:45 a.m.
6:15
p.m.
6:40
5:20 a.m.
p.m.
7:30
4:30 a.m.
p.m.
5:30 a.m.
6:30
p.m.
6:30
5:30 a.m.
p.m.
7:30
4:30 a.m.
p.m.
10
2.09
3.46
20
1.59
2.66
30
1.09
1.67
Aklo
Aklo
Coolik
Coolite
C.O.F.
Solex R
$With outside convos awnings tight against building on sides and top,
multiply over-all factor by I .4.
$Commerciol
shade bronze. Metal slots 0.05 inches wide, I7 per inch.
shade, oiuminum. Metal slats 0.057 inches wide, 17.5
$fWith
RADIATION
ABSORBED
Manufacturer
Blue Ridge Glass Corp.
Mississippi Glass Co.
Mississippi Glass Co.
Libbey-Owens-Ford
Pittsburgh Plate Glass
BY
HEAT
1
ABSORBING
GLASS
(%)
-IS&r
Color
Pole
Pale
Light
Light
Pale
Blue-Green
Blue-Green
Blue
Blue
Blue-Green
Pole Green
multicolor windows, ore the predominant color.
Radiation
Absorbed
56.6
69.7
58.4
70.4
48.2
50.9
Tnhlc 17 have been incrcxsetl to include the
I /.85 multiplier in Tfll~le 15.
tars in
GLASS BLOCK
Use of Table 17
- Solar Heat Gain Factors for Glass Block,
With and Without Shading Devices
The lactors in Tuhle I$ are used to dctcrminc the
solar heat gain thru all types ol’ glass block.
The transmission ol heat caused by a dilfercncc
between the inside and outdoor temperatures must
also be figured, using the appropriate “U” value,
Chpte~ 5.
Shading devices on the outdoor side of glass block
are almost as effective as with any other kind OC
g 1 ass since they keep the heat awry l’rom the g l a s s .
Shading devices on the inside are not effective in
reducing the heat gain because most of the heat
reHccted is absorbed in the glass block.
Example 6 - Peak Solar Heat Gain, Glass Block
Given:
Find:
I’eak solar heat gain
.
Solution:
By inspection of
Basis of Table 17
- Solar Heat Gain Factors for Glass Block,
With and Without Shading Devices
Table 15, the peak solar heat gain occurs
on July 23.
Solar heat gain
At 4:00 p.m. = (.39 X 164) + (.21 X 43) = 53
At 5:OO p.m. = (.39X 161) -I- (.21 X 98) = 84
At 6:00 p.m. = (.39 X 118) + (.21 X 1 4 4 ) = 76
The factors in Table I7 are the average of tests
conducted by the ASHAE on several types of glass
block.
Peak solar heat gain occurs at 5:00 p.m. on July 23.
Since glass block windows have no sash, the Eac-
.
TABLE II-SOLAR HEAT GAIN FACTORS FOR GLASS BLOCK
WITH AND WITHOUT SHADING DEVICES*
Apply Factors to Table 15
I
MULTIPLYING FACTORS FOR--..-GLASS "LVIR
n’npy
:
.,,.,
)., . ;:"
I
EXPOSURE
IN NORTH
LATITUDES
.'
‘9
.A,
I
I
.--,
I
..““,a
4.”
2”
.39
.35
74
.21
22
.27
.39
.24
.22
3.0
3.0
.-)_I?C
33
.--
,n
4.”
97
Northeast
East
Southeast
.-
3.0
3.0
e”“I..FYII
i -.;
C^..*Ls^r.
I
East
Northeast
.
:
.,. <:
“,.
:2,....
2,
North
South
Summart
Wintert
Southwest
West
Northwest
“Factors~include
,-
L A T I T U D E S ‘:’
‘.i ..-,. . .‘.Z-,
.
. . ,, , 2j ,z
.39
.27
I
I
correction for no rash with gloss block windows.
Equations
Solar heat gain without shading devices
= (Bi X Ii) + (6. X Ia)
Solar heat gain with outdoor shading devices
= (Bi X Ii + Ba X 1,) X .25
Solar heat gain with inside shading devices
= (Bi X Ii + Ba X Ial X .90
.
.21
.24
I
I
3.0
3.0
- -. . . . . . -. ,
Wintert
I
U--*L..,^-‘
. ..arl.ln~r,
West
..
Southwest
._ -!., i(
.
.
;
tllse the wmmer factors for. all latitudes, North or South. Use the
winter factor for intermediate seasons, 30’ to 50’ North 01 South
latitude.
Where:
Bi = Instantaneous transmission factor from Table 17.
B (I = Absorption transmission factor from Table 17.
= Solar heat gain value from Table 15 for the desired time and
Ii
wall facing.
42 = Solar heat gain value from Table 15 for 3 hours earlier than
Ii and same wall facing.
(:FIi\l’TI:l< 4. SOI.,\I<
111-,\‘I’
G.\IN
‘1‘111<11
l-5.5
(;L.:\SS
SHADING FROM REVEALS, OVERHANGS,
FINS AND ADJACENT BUILDINGS
i\II ~intl0~s
arc sliatlctl t o a greater o r Icrixr
tlcgrcc by tile projections close to it and by buildi n g s worii~tl i t . Tliis sll;itling retluccs the s o l a r
heat gain tltror~gh tllcse w i n d o w s by keeping t h e
direct rays oC the sun off part or all of the window.
T h e sllarletl p o r t i o n h a s o n l y t h e tliffusc c o m ponent striking it. Shading ol’ windows is signilic a n t i n nionuincntal type b u i l d i n g s w h e r e t h e
reveal may be large, even at the time of peak solar
heat gain. Clrtrrl I, this chapter, i s p r e s e n t e d t o
simplify the dotermination
of the shading of windows by these pro jcctions.
Basis of Chart 1
- Shading from Reveals,
Adjacent
Overhangs,
Fins
and
e location of the sun is defined by the solar
azimuth angle and the solar altitude angle as shown
in Fig. IS. The solar azimuth angle is the angle in a
horizontal plant between North and the vertical
plane passing through the sun and the point on ’
earth. The solar altitude angle is the angle in a
vertical plane between the sun and a horizontal
plane through a point on earth. The location of the
sun with respect to the particular wall facing is defined by the wall solar azimuth angle and the solar
altitude angle. The wall solar azimuth angle is the
angle in the horizontal plane between the perpendicular to the wall and the vertical plane passing
through the sun and the point on earth.
The shading of a window by a vertical projection
alongside the window (see Fig. 13) is the tangent of
the wall solar azimuth angle (B), times depth of
the projection. The shading of a window by a horiZC
31 projection above the window is the tangent
Or
qle (X), a resultant of the combined effects of
the altitude angle (A) and the wall solar azimuth
angle (Is), times the depth of the projection.
Tan X =
Tan A, solar altitude angle
Cos B, wall solar azimuth angle
T h e u p p e r p a r t o f Clln1.t 1 determines the tangent of the wall solar azimuth angle and the bottom
part determines tan X.
Use of Chart 1
- Shading from
Adjacent
FIG. 18 - SOl.hll
Buildings
Reveals,
Buildings
Overhangs,
Fins
and
The procedure to determine the top and side
shading from Chnrt I is.
1. Determine the solar azimuth and altitude angles l’roin Table 18.
‘\NGI.I<S
I
FIG.
19 -S HADING
BY
2. Locate the solar aLiniuth
upper part of Cl~clr.t /.
WA L L
PR O J E C T I O N S
angle on the scale in
3. I’roceetl hori/.on tally to the exposure clesirctl.
-I. Droll vertically to “Shading from Side” scale.
5 . ~Iultiply t h e d e p t h 0C the projection (plan
vielv) by the “Shading from Side.”
6. Locate the solar altitude angle
iowcr part ol C/,nrt I.
on
the scale in
nlltil the “Shatli~lg tronl
7 . ;\love Iiori~ontally
Side” value (45 clcg. lines) tlctcrniinctl in Step
4 is intersected.
x. DI-01) vertically t o “ S h a d i n g I’rotn 7’01”’ Cronl
intersection.
hlultil>ly
tile d e p t h 0C tllc lxojection (cleva!I.
tioil view) by the “Slla(li~lg l’rom TO]).”
SUN’S
RAYS
PLAN
FIG. 21 - .%ADINC
O F bYEAL ANI) O V E R H A N G
SUN’S RAYS
.
\\\ ’
\L
\
\
‘,
\\
Length of I)uiltling in shack, L
= 8.5 - 15 - (.l x 75) = 62.5 It
Height
of l)uildinS in shatle, H = 100 - (75 X .i) = -17.5 ft
The air contlitionctl lxtiltling is -shatlctl tfl a height of 47.5
ft and 62.5 Lt along the face at 4:00 pm. on July 23.
Example 8 -
Shading of Window by Reveals
Given:
^, L
A steel casement wintlow
reveal.
on the west sicie
with an S-inch 1
Find:
Shading by the reveal at 2 p.m. on July 23, 40” North
Latitude.
Solution:
From Table 18,
ELEVATION
FIG. 20 -
SHADING
OF
IjUlLDlNG
RY I-\D,JACENT
solar azimuth-angle = 242”
solar altitucte angle = 57”
From Chart I, shading from side reveal = .G X 8 = 4.8 in.
shading from top reveal = 1.8 X S = 14.4 in.
RlJIL.DING
Example 9 -
‘example
’
7 - Shading of Building by Adjacent Building
Given:
13uiltlings located as shown in Fig. 20.
Fintl:
Sharling at 4 p.m., July 23, bof I)uilding to 1)~ air conditioned.
Solution:
It is recommentletl
fhat the hkriltling plans ancl elevations
be sketcheci to scale with approximate
location of the sun,
Lo enable the engineer to visualize the shntling contlitions.
From Ta6le 19, solar azimuth angle = 265”
solar altitutle angle = 45”
From C/ICZV!
I, shading from sicte = ,I ft/ft a
shacting from top = .7 It / ft ,
Shading of Window by Overhang
and
Reveal
Given:
The same window as in Esarnple
6 inches above the window.
9 with a 2 ft overhang
Finch:
Shading by reveal and overhang at 2 p.m. on July
North
Latitude.
Solution:
Refer to Fig. 21.
Shading from sitle reveal (same as Exaurple
23, 40”
8) = 4.8 in.
Shatling from overhang = 1.8 X’ (24 i- 8) = 57.6 in.
Since the overhang is 6 inches above the wintlow,
of window shaded = 57.6 - 6.0 = 5 1.6 in.
the portion
CM,-\I”l‘l<R
,I, 401..\K III,:\ I
’
(i\IS
I’IIl<li
l-57
(i1.\5S
CHART 1 - SHADING FROM REVEALS, OVERHANGS, FINS AND ADJACENT BUILDINGS
Given:
Fintl:
Shading I)y rcvenl and overhang nt 2 pm, July 23,
40” North Latitutlc.
Solution:
From Table 18,
.\zimuth nnglc = 212” ’
Altitude nnglc = 57”
From Clrnrt I,
1. Enter at solar nrimuth angle (242”) to west
(it’) exposure shading from side = 0.G inch/
inch.
2. Enter at solar altitutlc angle (.57”) to shxling from sick (0.G inch/inch).. Shading from
top = 1.8 inch/inch.
3. Shading by reveal = 0.G X 8 = 4.8 in.
4. Shading I~yoverhnng=1.8(24+8)-6=51.6
in.
-I
DE (INCH/II NCI
\
\
\
t \
\
\,
‘\
30
\
35
40
45
50
55
60
65
+
70
75
8 0
,I
.I5
.2
.3
.4
3
.6
.7 .8
I:5
2
3
“SHADING’ FROM TOP ( I N C H / I N C H )
4
5
678
IO
15
20
l-58
,I’ART 1 . LO:\D ESTIbl.\‘I‘iNG
TABLE 18-SOLAR ALTITUDE AND AZIMUTH ANGLES
NORTH*
LATITUDE
LAT
0
0.
SUN
TIME
Jai n . 2 1
cir AZ
-
Fel b.
Alt i
6AM
7
a
9
IO
II
12 N
I PM
2
3
4
5
6
14
28
42
54
b5
70
as
54
42
28
14
15
30
44
58
71
79
71
58
44
33
I5
bAM
7
:
IO
II
I2 N
I PM
2
3
4
5
6
LA1
IO”
bAM
7
a
9
IO
II
I2 N
I PM
2
LAT
20”
:
5
6
6Alu
7
a
9
IO
II
l2N
I PM
LAT
30”
:
4
5
6
LAT
40’
bAt.
7
8
9
IO
II
I2 N
I PM
2
3
4
5
b
LAT
50”
6AM
7
a
9
IO
II
12 N
I PM
2
3
4
5
6
SOUTH*
LATITUDE
‘Use
months
SUN
TIME
indicated
II
13
I7
26
44
80
‘lb
!34
!43
,47
249
IO
24
37
48
57
60
57
48
37
24
IO
113
II7
124
I36
I55
I80
205
224
236
243
247
12
27
41
54
64
69
64
54
41
27
I2
7
103
I5
108
30
I I5 44
125
59
1 4 4 72
180
80
Tii 72
59
235
245 44
252
30
I5
257
- -I
114
121
I30
142
I58
I80
202
218
230
239
246
IO
23
36
47
55
59
5:
47
3t
23
IC,
106
I I2
121
133
I52
I80
2oE
227
235
246
254
95
1:
IO1
42
I08
55
I20
143
bC
JC I I80
c> 217
5 5 240
4i
252
2EI 259
I4
265
-
-
2
I4
24
3;
3E
4(
It
3;
2’
II
i
I<
3C
4C
4t
4(
4t
4(
3C
I<
;
125
I36
I49
lb4
I80
196
211
224
235
j I I(
I tI
II5
22 $
I3
3 :2 l4!
3. 7
lb:
3’?
l8(
3;7 rsr
3 :2 21!
2. 4 22’
I!5
!J 2,
f
1;
2.
2f
3(
-5
2z
I.
I
- -
3
IO
I5
:;
19
I5
IO
3
I25
138
I51
lb5
I80
195
209
222
235
IO
17
23
27
29
27
23
17
I C,
121
134
I48
lb4
1.20
196
212
226
239
IO
I9
27
34
39
4c I
%
34
27
I9
IC I
- luly 2 3 pug. 2 4
at
top
for
North
T-99
I IO
122
138
157
I80
203
222
238
250
261
- -
-
90
92
95
99
IO6
122
180
238
254
261
265
268
270
-
IO1
I: I
97
I It
2t
I06
12;
3c ; I lb
141
4c I
130
15s
5; 7 ISI
1%
b( 1 I80
201 T 7 209
215
4 ’I 230
23:
3t I 244
2 4 ‘ 2( 254
25:
I: ; 263
115
124
134
146
lb2
I80
I98
214
226
236
245
Ma
iF
78
77
74
68
53
0
iii
!92
286
283
282
4
‘8
I2
;4
a5
‘0
&
j4
&2
18
14
54
36
0
124
106
297
293
29 I
7
81
83
84
84
84
0
7%
276
276
277
279
282
3
I7
32
46
50
73
30
73
60
46
32
I7
3
JO
72
72
72
67
53
0
307
293
288
288
288
290
4
I8
32
46
59
72
81
72
59
46
32
I8
4
19
84
89
94
IO2
117
l8C
Tz
256
261
271
2Jt
281
7
20
34
48
b2
75
90
75
62
48
a
I?
31
44
56
67
71
I
56
44
31
I9
b
-E
8;
7
I9
30
41
51
58
.L!
58
51
41
30
I9
7
-
a
15
90
I.5
30
89
30
45
8 9 44
b0
89 5 8
88
71
75
90 L
7 ,9
75
t-i ‘I
b0
;8
45
(4
30
IO<
T5 2 7 0 I5
-
6
I?
30
4c
47
50
4i
4c
3c
I9
6
02
03
06
II2
I27
80
GT
!48
!54
157
258
-
1
lb
31
46
51
75
39
75
bi
46
31
lb
2
-
Latitudes:
.
I
2
4
I5 102
30
I03
44
IO6
58 112
71
127
7’9
I80
71
233
58 248
44 254
30 257
15 258
4
III
‘8 II3
I2
117
;4 126
,5
144
‘0 180
,5 216
54 2 3 4
+2 243
28 247
I4 2 4 9
14
II4
27
117
41
122
53
131
62
I48
67 180
62 212
53 229
41
238
27 243
I4 2 4 6
bAM
7
IB
9
I( 0
II
I: ZN
I PM
2
3
4
5
6
9
II6
23 121
35 128
46
139
53
156
57
180
53 204
46 221
35 232
23 239
9 244
6AM
7
a
9
I0
I I
I 2N
I PM
2
3
4
5
6
-7
lb
il
46
jl
75
39
ii
ii
16
31
lb
2
78
81
83
84
84
84
0
276
276
276
277
279
282
I
I5
IO
$4
j?
72
30
72
j’?
14
30
15
I
90
92
95
99
106
122
I80
238
254
261
265
268
270
I2
103
27
108
41
II5
54
I25
64
144
69
I80
64 216
54 235
41
245
27 252
12 257
IO
24
37
48
57
10
57
48
37
24
IO
I13
117
124
I36
155
180
205
2 2 4
236
243
2 4 7
-7
75
79
82
85
88
0
272
275
278
281
285
289
8
68
21
72
35
75
48
77
62
77
76
74L
87
CI
76 28t
b2 28:;
4 8 2 8 :I
3 5 2%>
21
2% I
8 2% ,
T
I8
32
4b
59
72
81
i
55
4c
32
IE
4
79
84
89
94
102
I I7
I8C
243
258
2bt
271
276
281
14
28
42
55
66
JO
66
55
42
28
I4
95
IO1
108
12c
143
I80
217
24C
252
259
265
IO
23
36
47
55
59
55
47
36
23
IO
6
19
30
40
47
50
47
40
30
19
6
114
121
130
142
158
I80
202
218
230
239
246
To -77
79
23
35
86
48
93
61
103
73
122
it!? 180
7 3 218
bl 257
48 267
35 274
23 281
IO 288
II
6’I
24
71
37
8: :
49
8l i
62
9t j
75 II: 2
83
181I
75 241 3
62 26‘4
4 9 2 7 :2
37 2713
2 4 282
4
II
29 I
2
14
24
32
38
40
38
32
24
I4
2
II5
124 II
134 21
146 29
162
35
180 1 37
198
35
214 29
226 21
236
II
245
t
7
71
20
75
34
79
48
82
62
85
75
88
90
c
75 272
62 275
48 2 7 E
34 281
2 0 285
7 289
13 74
83
24
35
93
47 104
57
II8
66 143
J&c
?c
bb 217
5 7 242
47 256
3s 267
24 277
13 286
-
83
94
106
120
137
I57
l8C
203
223
24C
254
2bt
277
-
I5
25
34
44
52
58
60
58
52
44
34
25
15
-
I8
74
27
85
37
97
46
IIC,
55 I28
61
151
63E I
bl
209
55 232
4 6 2 5 CI
37 263I
2 7 276,
I8 28C
use
Dec. 22
Alt AZ
4
67
3
70
I8
68
I7
72
12
72
32
68
45
67
I6
72
58
61
50
67
73
53
JO
44
77
0
I 30
0
J O 316 73 307
58 299
( 50 293
45 293
16 2 8 8
32 292
32 288
I8 2 9 2
I7 288
4 293
3 290
I
9
I 0:
I I:
12’
15
I 81
55
23
24’
251
265
279
-
9:
90
Nov. 2 I
I4lt A Z
;-.-i-i
>O
89
‘5
88
‘0
0
‘5 272
,O 2 7 1
t5 2 7 1
IO 2 7 1
I5 270
t
1
t
c
77
88
100
I I4
I31
152
180
208
229
246
260
272
283
-
Jov. 21
months
at
Dec. 22
bottom
for
5
IO6
112
121
I33
152
I8C
20E
227
235
24E
254
IO
72
23
75
35
81
48
93
bl
IO:
73
12;
80
I8(
73 238
bl
25;
48 26;
35 27’
23 281
IO 285
I5
25
34
44
52
58
60
58
52
44
34
25
15
77
88
I00
II4
I31
I52
I80
208
229
246
260
272
283
. Ian. 21
South
9
18
28
37
44
49
51
49
44
37
28
I8
9
83
94
IO6
120
137
157
I8C
203
223
240
254
266
277
:ab. 20
Latitudes.
12
23
33
42
48
50
48
42
33
23
I2
99
II0
122
138
157
I80
203
222
238
250
261
IO
I9
27
34
39
40
39
34
27
I?
IO
IO1
II4
127
143
lb0
I80
200
217
233
246
259
dar. 22
T--
Oct. 23
Alt AZ
78
77
74
68
53
0
307
292
286
283
282
7 -z
IF
8i
31
9E
4 4 104
51
I Ii
6;
I4(
71
I 8(
s
22(
24:
5t
4’
2%
31
2bt
27:
I(
28(
-i t
7
81
I5
7: 2
I3
7’
26
8( 3
24
8:
I?
?I
10;
37
8’ ? 35
9 : 30
41
I I:
49
I01 3
47
IO‘
60
II.4
5 7 v/t 51
12’
66 14:
58
15
69
1319
l8( blI8(
73 I81 3 _ 70
58 20’
69 22:2
51. 23l
60 24, b
if 2
4:
49 261 3
47 251
41
24
37 27 I
35 26:
30 251
7
2 6 28L
2 4 27; I9 26’.
I5 2 8 8
I3 286
7 279
>ct. 23
and
4r
iii
-
i
5
,O
.4
,8
.I
‘9
T
i8
14
IO
i5
2
7
IO‘
Iii
I4(
18(
zi
24:
251
26!
27:
28(
9
I8
28
37
44
49
51
49
44
37
28
I8
9
- IO1
II4
127
143
I60
I80
200
217
233
246
259
;ept. 22
June 2 I
20
F
-
5
I5
24
32
37
39
37
32
24
I5
5
IO
I7
23
27
29
27
23
I7
IO
III I
II! I
13 I
l4! j
lb: 2
18( JI91 3
2l! 5
22’ I
24 I
25L7
I21
I,34
148
lb4
I8C ,
I96
212
226
2,39
Apr. 20
8
17
24
28
30
28
24
I7
8
3
IO
15
19
20
I?
15
IO
3
125
136
I49
lb4
I80
I96
211
224
235
I25
I38
I51
lb5
I80
I95
209
222
235
lJuna 21
May
5
17
28
38
44
47
44
38
28
17
5
5
I4
21
25
27
25
21
14
5
SUN
TIME
bAM
7
117
124
133
I45
163
I80
197
215
227
236
243 .
:
I0
I I
I2 N
I PM
2
3
126
136
149
lb4
I80
196
211
224
234
6AM
7
a
9
IO
II
I2 N
I PM
2
3
4
5
6
127
138
I51
lb5
I80
195
209
222
233
:
6
bAk.4
7
a
1’0
II
12 N
I PM
2
3
4
5
b
6AM
7
a
6
12
15
17
I5
I2
6
I39
152
lb6
I80
194
208
221
21
9
0
I
2N
I PM
2
3
4
5
6
SUN
TIME
.
l-59
CHAPTER 5. HEAT AND WATER VAPOR FLOW
THRU STRUCTURES
This chapter presents the methods and data for
determining the sensible and latent heat gain or loss
thru the outdoor structures of a building or thru a
structure surrounding a space within the building.
It also presents data for determining and preventing
water vapor condensation on the enclosure surfaces
or within the structure materials.
Heat flows from one point to another whenever a
t
>eraturc clifference exists between the two points;
the direction of flow is always towards the lower temperature. Water vapor also flows from one point to
another whenever a difference in vapor pressure
exists between the two points; the direction of flow is
towards the point of low vapor pressure. The rate at
which the heat or water vapor will flow varies with
the resistance to flow between the two points in the
material. If the temperature and vapor pressure of
the water vapor correspond to saturation conditions
at any point, condensation occurs.
HEAT
FLOW
THRU
BUILDING
‘I = UAAt,
where
q = heat flow, Btu/hr
U = transmission coefficient,
Btu/(hr)(sq ft)(dcg F temp cliff)
A = arca of surface, sq ft
At, = equiv temp diff F
Heat loss tllru the exterior construction (walls and
is normally calculated at the time of greatest
bent /loru. This occurs early in the morning after a
roof)
few hours of very low outdoor temperatures. This
approaches steady state heat flow conditions, and for
all practical purposes may be assumed as such.
Heat flow thru the interior construction (floors, I
ceilings and partitions) is caused by a diflerence in
temperature of the air on both sides of the structure.
This temperature difference is essentially constant
thruout the day and, therefore, the heat flow can be
determined from the steady state heat flow equation,
using the actual temperatures on either side.
STRUCTURES
Heat gain thu the exterior construction (walls
and roof) is normally cnlculnted at the time of
greatest heat /?OZO. It is caused by solar heat being
absorbed at the exterior surface and by the temperature difference between the outdoor and indoor
a:-- Both heat sources are highly variable thruout
2,
one day and, therefore, result in unsteady state
heat flow thru the exterior construction. This unsteady state flow is difficult to evaluate for each individual situation; however, it can be handled best
by means of an equivalent temperature difference
across the structure.
, The equivalent temperature difference is that temperature difference which results in the total heat
flow thru the structure as caused by the variable
solar racliation and outdoor temperature. The equivalent temperature difference across the structure
must take into account the different types of construction and exposures, time of day, location of the
building (latitude), and design conditions. The heat
flow thru the structure may then be calculated, using
the steady state heat flow equation with the equiv--- --.__
alent temperature difference.
- il
EQUIVALENT TEMPERATURE DIFFERENCE SUNLIT AND SHADED WALLS AND ROOFS
The process of transferring heat thru a wall under
indicated unsteady state conditions may be visualized by picturing a 12-inch brick wall sliced into
12 one-inch sections. Assume that temperatures in
each slice are all equal at the beginning, and that the
indoor and outdoor temperatures remain constant.
When the sun shines on this wall, most of the solar
heat is absorbed in the first slice, Fig. 22. This raises
the temperature of the first slice above that of the
outdoor air and the second slice, causing heat to
flow to the outdoor air and also to the second slice,
Fig. 23. The amount of heat flowing in either direction depends on the resistance to heat flow within
the wall and thru the outdoor air film. The heat flow
into the second slice, in turn, raises its temperature,
causing heat to flow into the third slice, Fig. 24.
This process of absorbing heat and passing some on
to the next slice continues thru the wall to the last
or 12th slice where the remaining heat is transferred
to the inside by convection and radiation. For this
particular wall, it takes approximately 7 hours for
FIG. 22 - SOL,\R HEAT ARSORBED
IN
FIRST S LICE
FIG. 25
- BEIIAVIOR
SECOND
T IME
HEAT
01; ABSORBED
INTERVAL
AUSORBED
PLUS
DURING
S OLAR HEAT
AD D I T I O N A L
DURING
S OLAR
T HIS INTERVAL
’
.
- B EHAVIOR
FIG. 23
DURING
OF
S ECOND
ABSORBED S OLAR HEAT
T IME
INTERVAL
3
FIG.
-
FIG. 24
- B EHAVIOR
DURING
OF
T HIRD
ABSORI~ED
TIME
S OLAR HEAT
INTERVAL
solar heat to pass thru the wall into the room.
Because each slice must absorb some heat before
passing it on, the magnitude of heat released to
inside space would be reduced to about 10% of that
absorbed in the slice exposed to the sun.
These diagrams do not account f o r possible
changes in soln~ intensity or outdoor temperature.
26
T HIRD
-B EHAVIOR
OF
ABSORBED S OLAR HEAT
T IME INTERVAL PLUS ADDITIONAL S OLAR
ABSORBED DURING T HIS INTERVAL
DURING
HEAT
The solar heat absorbed at each time interval by
the outdoor surface of the wall throughout the day
goes thru this same process. Figs. 25 and 26 show the
total solar heat flow during the second and third
time intervals.
A rise in outdoor temperature reduces the amount
of absorbed heat going to the outdoors and more
flows thru the wall.
This same process occurs with any type of wall
construction to a greater or lesser degree, depending
on the resistance to heat how thru the wall and the
thermal capacity of the wall.
NOI‘E:
The thermal capacity of a wall or roof is
the density of the material in the wall or
rool’, times the specific heat of the material,
times the voli~n~c.
This progression of heat gain to the interior may
occur over the full 24-hour period, and may result
in a heat gain to the space during the night. If the
equipment is operated less than 24 hours, i.e. either
skipping the peak load requirement or as a routine
procedure, the nighttime radiation to the sky and
the lowering of the outdoor temperature may decrease the transmission gain and often may reverse
it. Therefore, the heat gain estimate (sun and transmission thru the roof and outdoor walls), even with
equipment operating less than 24 hours, may be
evaluated by the use of the equivalent temperature
data presented in Tables 17 and 20.
Basis of Tables 19 and 20
- Equivalent Temperature Difference for Sunlit
Shaded Wails and Roofs
and
tions using Schmidt’s method based on the following
conditions:
1. Solar heat in July at 40” North latitude.
2. Outdoor daily range of dry-bulb temperatures,
20 deg F.
3 . hIaximum outdoor temperature of195 F db and
a design indoor temperature of 80 F db, i.e. a
design difference of 15 deg F.
4 . Dark color walls and roofs with absorptivity
of 0.90. For light color, absorptivity is 0.50;
for medium color, 0.70.
5 . Sun time.
3e specific heat of most construction materials
is &ipproximately 0.20 Btu/(lb)(deg F); the thermal
capacity of typical walls or roofs is proportional to
the weight per sq ft; this permits easy interpolation.
for
Sunlit
Civcn:
.\ flat roc,f cxposc’l to the sun, w i t h I)uilt-up roofing, 1%
in. insulation, 3 in. woo(l tlcck anti suspentlctl acoustical
tile ceiling.
Room design tcmI)erature = X0 1: tll)
Olittloor tlcsign tcmpfratul-e
= 95 F tll)
I)aily range ‘= 20 tleg F
Find:
Equivalent temperature tlilfercncc
at 4 p.m. July.
Solution:
avt/sq ft = 8 + 2 + 2 = 12 Il,/sq ft (Table -77. @ge i/i
Equivalent temperature
difference
= 43 tleg I; (Table 20, intcrpolatetl)
Example 2 - Daily Range and Design Temperature
Difference
Correction
At times the daily range may be more or less than 20 deg F;
the difference between outdoor ancl room clesign temperatures
may he more or less than 15 deg F. The corrections to be
applied to the equivalent temperature cliff rence for combinaf
tions of these two variables are listed in the notes following
Tables 19 and 20.
Tubles I9 nnd 20 are analogue computer calcula-
Use of Tables 19 and 20
- Equivalent Temperature Difference
Shaded Walls and Roofs
Example J - Equivalent Jemperafure Difference, Roof
and
The equivalent temperature differences in Tables
19 and 20 are multiplied by the transmission coefFicients listed in Tables 21 thru 33 to determine the
heat gain thru walls and roofs per sq ft of area during the summer. The total weight per sq ft of walls
and roofs is obtained by adding the weights per sq
ft of each component of a given structure. These
weights are shown in italics and parentheses in
Tables 21 thou 33.
Given:
The same roof as in Exa~r~/Ae I
Room design temperature = 78 F dl)
Outdoor design temperature = 95 F tlh
Daily range = 26 cleg F
Find:
Equivalent
tempqrature
difference
I
under
changed
conditions
Solution:
Design temperature ciifference = 17 deg F
Daily range = 21.3 deg F
Correction
to
eqttivalent
temperature
difference
= -1 cleg F (TaDle 20A, interpolated)
Equivalent temperature difference = 43 - 1 = 42 deg F
Example 3 - Other Months and Latitudes
Occasionally the heat gain thru a wall or roof must be known
for months and latitudes other than those listed in Note 3
following Table 20. This equivalent temperature difference is
determined from the equation in Note 3. This equation adjusts the equivalent temperature difference for solar radiation
only. Additional correction may have to be made for differences Ijetween outdoor and indoor design temperatures other
than 15 cleg F. Refer to Tables 19 and 20, pages 62 and 63,
ancl to the correction TaBle 20A. Corrections for these differences must Ile made first: then the corrected equivalent
temperature differences for both sun and shade must be
applied in corrections for latitude.
Given:
12 in. common brick wall facing west, with no interior
finish, located in New Orleans, 30” North latitude.
Find:
Equivalent
temperature
difference
in
November
at
12
noon.
Solution:
The correction for design temperature difference is as
follows:
‘J
Outtloor tlcsign tlry-l)ull)
=!)r,-ir,=XOF
lw~pcr;~turc in Novcrnlvx at 3 p.m.
A’,~,,,
for west wall, in sun
= i (Tmhlr 1 9 ) - I I .5 = - 4,s tlcg 1;
TABLE 19-EQUIVALENT TEMPERATURE DIFFERENCE (DEG F)
FOR DARK COLOREDt, SUNLIT AND SHADED WALLS*
/
or/
-Jq
Based on Dark Colored Walls;b“;“F db Outdoor Design Temp; Constant 80 F db Room Temp;
20 deg F Daily Range; 24-hour Operation; July and 40” N. Lat.i
SqN
WEIGHT
OF WALLf
( I b / s q ft)
EXPOSURE
TIME
I
a /9
110 I 11 I 12
i
101
South
%
71
61
31
2
21
Southwest
31
41.5
- 2 - 3 -4 -2
2
1
O - l
6
6
5
5
9
8
7
7
0 -1
4
9
8
13 13
-2 -3 -3
3
110
.7
7
6
12 12 12
0 -1
5
4
9
9
11 1 1
-1 -2 -2
3
3
2
8
8
7
10
10
9
11
2
8
10
1
5
10
8
0
1
7
9
1
4
9
8
1
6
9
0
4
9
8
0 -1
O-1
6
5
8
7
-1
3
8
8
-1
3
7
8
1
0
0 -1 -1
5
4
3
3
2
12
11
10
9
8
20 18 16 15 13
west
Y
11
4
7
9
East
Y-4
0I I i I
A M
PM
Northeast
Southeast
Northwest
6
6
6
-2
1
-;;-:,
0
)
61
7)
81 4
IO
0
6
4
LJ
;,
0
11
A M
6
a 10
;
6
6
12. 14
;
Jl;
0
0
1
"12
1
2
2
3
7
8
5
6
7
PM
I
,
9
13 12
10
1; 12
12
5. 5
3
4
5
4
I
I
I
.
SUN TIME
E q u a t i o n H: e a t G o i n T h r u W a l l s , B t u / h r= ( A r e a , s q f t ) x ( e q u i v a l e n t t e m p d i f f ) x ( t r a n s m i s s i o n c o e f f i c i e n t U , T a b l e s 2 1 thru 25)
* A l lv a l u eosr e f o r b o t h i n s u l a t e d a n d u n i n s u l a t e d w a l l s .
i F o ro t h e r c o n d i t i o n s , r e f e r t o c o r r e c t i o n p
s ao g
n e6 4 .
f " W e i g h t p e r s q f t " v a l u e s f o r c o m m o n t y p e s o f c o n s t r u c t i o n o r e l i s t e d i n Tables 21
thru25.
F o r w a l lc o n s : r u c t i o n sl e s s t h a n 2 0 I b / s qi t , u s e l i s t e d v a l u e s o f 2 0 I b / s qf t .
‘1
!. i
\
C:l~i\I’~I‘El<
.5. HE.\‘I‘
,\NI) W.\~I‘I:K
\‘.\I’OR
Ipl.OW
‘1.1~IRlJ
l-63
S’I‘lIII(:‘l‘IJl<~S
TABLE 20-EQUIVALENT TEMPERATURE DIFFERENCE (DEG F)
‘71.
FOR DARK COLOREDi, SUNLIT AND SHADED ROOFS*
Based on 95 F db Outdoor Design Temp; Constant 80 F db Room Temp; 20 deg F Daily Range;
24-hour Operation; July and 40’ N. Lat.t
CONDITION
rosed
to
SUn
WEIGHT
OF ROOFS
(lb/v
ft)
10
20
40
60
80
Covered
with
Water
Sprayed
SUN TIME
I
6
A
f
,
M
.\
AM
PM
10
11
12
1
2
3
4
j
6
7
8
9
1
2
- 4 -6 - 7 -5 -1
0 - 1 - 2 - 1
2
4
3
2
3
6
9
8
6
7
8
7
9
15
24
32
38
43
46
45
41
35
20
22
16
10
7
3
16
23
30
36
41
43
43
40
35
30
25
20
16
16
I6
23
22
22
28
27
26
33
31
28
38
35
3_2
40
3%
35
41
39
37
39
38
37
35
36
35
32
34
34
28
31
34
24
28
32
15
20
25
30
12
17
22
27
8
13
18
23
19
13
22
15
20
15
18
16
16
15
14
15
12
14
10
12
6
10
2
16
10
5
10, 1 2
14
15
16
15
14
12
2
7
10
1
5
8
1 - 1 - 2 - 3 - 4 - 5
3
1 -1 - 2 - 3 - 3
6
4
3
2
1
0
I5
9
5
18
17
16
15
14
12
10
6
2
1
13
14
14
14
14
13
12
9
7
5
0 - 1 - 2 - 2 - 3 - 3
3
1
0
0 - 1 - 1
8
10
12
13
14
13
12
11
10
8
6
6
7
a
13
11
12
10
11
13
2
4
10
-I -I
0
-2,-2 - 2 - 2
5
12
20
40
60
- 5 -2
- 3 - 2
20
40
60
- 4 -2
- 2 -2
- 1 - 2
-I
Shaded
9
11
0
0
2
4
8
12
- 1
-I
0
2
5
- 2 - 2 - 2
0
2
7
lop12
4
3
4
5
1 - 1 - 3
6
4
2
11
9
6
16
13 11
20 18 14
2
1
0 - 1
3
4
”
6
7
8
9
10
11
12
1
2
3
4
AM
5
9
10
11
12
1
2
A
PM
M
5:
:-..I
SUN TIME
Equation: Heat Gain Thru Roofs,
sq ft) X (equivalent temp diff) X (t ransmission
Btu/hr = (Area,
*With attic ventilated and ceiling insulated roofs, reduce equivalent temp diff
For peaked roofs, use the roof oreo projected on CI horizontal plane.
coefficient U, Tables 27 or 28)
25yo.
tFor other conditions, refer to corrections below and on page 64.
:“Weight per sq ft” values for common types of construction are listed in Tables 27 or 28.
TABLE
20A-CORRECTIONS
TO EQUIVALENT TEMPERATURES (DEG F)
OUTDOOR
FO:E%NNTH/’
AT 3 P.M.
-/MINUS
ROOM TEMP
(deg F)
-30
-20
-10
0
5
10
15
20
25
30
35
40
DAILY RANGE (deg F)
12
14
1‘6
-41
-42
-31
-32
-2.L
-22
Lll J-12
la
20
22
24
26
-a3 , - 4 4
-33 -34
- e -24
-14
-,;3
-45
-35
-25
-46
-36
-26
- 1 5
-16
-47
-37
-27
-17
-48
-38
-28
-18
2;.
--;I,
-12
- 7
- 2
3
8
13
18
23
-4’
.
-’
- 1
4
9
14
19
24
29
‘1’2
_Tgr
\-
3.:
B-;,.7
13
ia
23
28
2”,
12
17
22
27
;J+
‘Ll
‘b
11
16
21
26
0‘
3”
- 1
4
10
9
15
20
25
14
19
24
28
30
32
34
-50
-40
-30
-20
-51
-41
-31
-21
-52
-42
-32
-22
36
-53
-43
-33
-23
-13
-8
- 3
2
-15
-10
-5
0
-16
-17
-12
- 7
- 2
-ia
-13
- a
- 3
7
12
17
22
5
10
15
20
4
9
14
19
3
8
13
18
2
7
12
17
-
-
4
3
2
1
9
9
9
9
6
11
16
21
-11
-6
- 1
38
-
5
4
3
2
4
4
4
4
1
6
11
16
-55
-45
-35
-25
1-64
I’AKT
Corrections
to
Equivalent
Temperature
Differences
in
Tables 19 & 20 for Conditions Other Than Basis of Table
I
.
LO.\D ES-I‘lMr\-I‘ING
Al,,*
cliffercncc for same wall or
= cclilivalcnt tcmpcrattirc
roof in shade at clesirctl time of (lay. corrcctecl if
neccsqary for design conditions.
3 ,‘,I&
= eclilivalent tcmperatnrc
ciilFcrcnce for wall or roof
cxl~~secl to the sun for the clesirccl I ime of clay. carrcctccl if ncccssary for tlesign conditions.
Note: Light color = white, cream, etc.
;\Ieclilim color = light green, light blue, gray, etc.
I)ark color = Clark Ijlue, dark red, dark brown, etc.
5. Other Iatitllcle, other month. light or metlicim color walls
or roof.
The coml,inecl formulae arc:
Light color walls or roof
2. Shaclccl walls
For sha~lccl wa1l.s on any cxpc~surc, tlse the values of equivalent temperature clifferencc listed for north (shacle), carrcctecl if necessary as shown in Correction I.
3. Latitudes other than 40” North and for other months with
cllfferent solar intensitlcs. Tables 19 mtl 20 values are~approximately correct for the east or west wall in any latitude(luring
the hottest weather..- In lower latituclcs when the
maximum solar altitude is 80” to 90” (the maximum occurs
at noon), the temperature difference for either south or
north wall is approximately the same as a north or shade
wall. See Table 18 for solar altitude angles,
Meclium color walls or roof.
R
At,, = .78 j{f 1tc,,,
Tab/r J9:
Northeast
East
Southeast
South
Southwest
West
Northwest
North (shade)
= equivalent temperature difference for month and
time of clay desired.
equivalent temperature difference for same wall or
roof in shade at desired time of clay, corrected if
necessary for design conclitions.
4t
equivalent temperature difference for wall or roof
En1 exposed to the sun for desired time of clay,-corrected if necessary for design conclitions.
-\
sRs! = maximum solar heat gain in I%tu/(hr)(sq ft) thru
glass for wall facing or horizontal for roofs, for
month and latitude desired, Table 15, page q-1, or
7
Table 6, l-‘age 29.
ft) thru
R,,:,
= maximum solar heat gain in Btn/(hr)(sq
glass for wall facing or horizontal for roofs, for
July at 40” North latitude, Tnble 15. page ff, or
=
Table 6, jmge 39.
.Qaffrfile
3 illustrates
the procedure.
30
ltc = ltcs -f To (At< ,,,, - -\t(,J
hfeclium
TRANSMISSION COEFFICIENT U
thru
structure. The reciprocal 0E
building
structure
t o t a l rcsist:lnce
heat How
thru the
the U value for any wall
is the total resistance of this wall
‘Phe
a
ft)(deg F temp cliff). T h e r a t e t i m e s
How of heat.
the
to the
of any wall to heat flow is
s u m m a t i o n o f t h e r e s i s t a n c e i n e a c h c o m p o n e n t 0E
the
structure
and
surface
in
Tnb1e.s
most
the
films.
common
resistances
The
-31 tlcrzl
tyllcs
of
the
transmission
outdoor
and
coefficients
33 have been calculated for
of construction.
Basis of Tables 21 thru 33
- Transmission Coefficients U for Walls, Roofs, Partitions,
Ceilings, Floors, Doors, and Windows
= 55 4tc,,L + .45 4t,,*
Tnbies -71 that 3 3 c o n t a i n c a l c u l a t e d U v a l u e s
.70
4tca
based on the resistance
The
where:
roof
transferred
the tcmperaturc difference is the
listed
4tc = At(,n + 3 (4\t(,,,, - 4to) = .i8 4tc,,; + .22
or
heat is
i n Ktu/(hr)(scl
the
color wall or roof:
= equivalent
or U value is the rate at
Transmission coefficient
which
inside
4. Light or meclium color wall or roof
Light color wall or roof:
4tc
Use Exposure Value
Southeast
East
Northeast
North (shade)
Northwest
West
+
Southwest
South
.hutJ2 IA titlltle
where
4tcr
.
6. I;or South latitudes, use the following exposure values from
The temperature differential Ate for any wall facing or roof
and for any latitude for any month is approximatecl
as
follows:
4te
R
+ (I - .78 2) -Itcs
!!I
temperature
desired.
difference
for
color
of
wall
for
listed in
Tnble
resistance of the outdoor surfxc
and winter conditions and the inside
film is listed in Table 3f.
summer
~surface
34, j3nge 78.
film coefficient
Note:
Tlrc tliffercrrcc ljetwecn summer and winter,
tr;rrlsmissiori cocffkients for ;I typical wall
is ricgligil)le. For example, with ;I trnnsrrrission coefficient of 0.3 I~tu/(hr)(sq ft) (F)
_ [or w i n t e r contlitions, t.lic c o e f f i c i e n t f o r
summer conditions will be:
I. Thcr-mal rcsistancc
zz
R
(winter) ol
twll
)
Example
4
-
Transmission
Coefficients
Civcn:
brasonry partition matlc of 8 in. I~ollow clay tile, Itoth sides
finished, metal lath plasterctl
on furring with ‘% in. sand
plaster.
Find:
Transmission
coelfirient
Solution:
Transmission coefficient U
= 0.18 I%tu/(hr)(sq ft)(tleg F), Table 26, /xtge 70
1
I
-=
- = 3.33
0.3
I/
2. Cktdoor film therm;\1 resistance (winter)
= 0.17 (Tnble 34)
3. Thermal resistance of wall without outdoor air
film (wiritcr) = 3.33 - 0.17 = 3.16
-1. Outdoor film thermal rcsist:ince
= 0.25 (Table 34)
6. Transmission coefficient U of wall in summer
1
1
=----_
= 0.294
R
3.41
7. Difference between summer- and winter transmission becomes greater with larger U values
and less with smaller U values.
Ceilings,
Floors,
Doors,
U
and
for
Wails,
5
- Transmission
Insulation
Coefficient,
Addition
of
The transmission coefficients listed in Tables 21 thru 30 do not
include insulation (except for flat roofs, Table 27, page 71).
5. Thcrrnal rcsistancc of wall with outdoor air
film (surnrncr) = 3.16 + 0.25 = 3.41
Use of T,ables 21 thru 33
- Transmission Coefficients
Example
(summer)
Roofs,
Partitions,
Windows
The transmission coefficients may be used for calculating the heat flow for both summer and winter
conditions for the average application.
Frequently, fibrous insulation or reflective insulation is included in the exterior I)uilding
structure. The transmission
coefficient for the typical constructions listed in Tables 21 thru
30, with insulation, may Ire determined from Tabfe3Z,page 75.
Given:
Masonry wall consisting of 4 in. face brick, 8 in. concrete
\cinder block, metal lath plastered on furring with sh in. ’
sand plaster and 3 in. of fibrous insulation in the stud space.
Find:
Transmission
coefficient.
Solution:
Refer to Tables 22 and 31.
C’ value for wall without insulation
= 0.24 Btu/(hr)(sq ft)(dcg F)
CT value for wall with insulation
= 0.07 I%tu/(hr)(sq ft)(deg F)
l-66
\
I’,\RT I . LOI\I) ESl‘I~I.\‘I‘IN(;
+ TABLE 2 1 -TRANSMISSION COEFFICIENT U-MASONRY WALLS*
FOR SUMMER AND WINTER
Btu/(hr) (sq ft) (deg F temp diff)
All numbers in parentheses indicate weight per
sq f t . T o t a l w e i g h t p e r s q f t i s s u m o f w a l l a n d f i n i s h e s .
INTERIOR
%”
IHICKNESS
inches)
and
YElGHT
(lb p e r
sq ftl
G iypsum
%*
Board
Plaster
on Well
NOW
( Plaster
IBoard)
(21
T
gq-!g
SOLID BRICK
Face 8
Common
8 (87)
2 (123)
I6 (173)
.48
2.5
.27”
.41
.31
.25
.45
.33
.26
.41
.30
.25
Common
Only.
a (80).
I2 (1201
16 (1601
.41
.31
.25
.36
.2a
.23
.39
20
.24
.35
.27
.23
(100:
(15-o:
(200’
(300’
.67
.55
.47
.36
.55
.47
.41
.32
.63
.52
.45
.35
8 (26)
12 (40)
.34
.25
.30
.23
6
8
10
12
(70)
(93)
(117
(140
.75
.67
.61
.55
6 140)
8 (53)
10 (66)
12 (80)
6 (15)
8 (20)
FINISH
I/s If
Metal
Lath
Plastered
on Furring
=h/4n
Sand
P‘laster(l)
Yin
1t wt
Plaster(J)
Insulating
Board
Plain or
Plastered
on Furring
Gypsum or
Wood Lath
Plastered
on Furring
l/i ‘I
t/$ #
Sand
1t wt
Plaster(7J
“:,
Plaster(Z)
.31
.25
.21
.28
.23
.I9
.29
.23
.20
.27
.22
.I9
.22
.19
.I6
.16
.I4
.I3
.28
.23
.19
.26
.?2
.I8
.26
.22
.18
.25
.21
.I8
.21
.18
.16
.I5
.14
.I2
.53
.46
.40
.32
.39
.34
.31
.26
.34
.31
.28
.24
.35
.31
.2a
.24
.32
.29
.27
.23
.26
.24
.22
.19
.I8
.17
.I6
.I5
.32
.24
.30
.23
.25
.20
.23
.18
.23
.I8
.22
.18
18
.15.
.12
.I4
.55
.49
.44
.40
.69
.63
.57
.52
.58
.53
.49.
.45
.41
.39
.36
.34
.36
.34
.32
.31
.37
.35
.33
.31
.34
.32
.31
.29
.27
.26
.25
.24
.18
.17
.17
.16
.31
.25
.21
.18
.2a
.23
.19
.17
.30
.24
.20
.I7
.27
.23
.19
.I5
.23
.19
.17
.15
.21
.I8
.I6
.14
.22
.18
.I5
.I4
.21
.I8
.14
.14
.I8
.16
.14
.I2
.14
.12
.I 1
.lO
12 (30)
.13
.lO
.08
.07
.13
.lO
.OE
.07
.13
.I0
.08
.07
.13
.I0
.08
.07
.I2
.09
.08
.07
.I 1
.09
.07
.07
.I 1
.09
.08
.06
.I 1
.09
.07
.06
.13
.lO
.08
.07
.09
.07
.06
.06
8 (43)
12 (63)
.52
.47
.44
.41
.48
.45
.43
.40
.23
.22
.17
.16
a (37)
12 (53)
8 (32)
12 (43)
.39
.36
.35
.33
.37
.35
.34
.32
.20
.19
.15
.I5
.35
.32
.32
.29
.34
.31
.31
.28
.19
.18
.15
.14
8 (39)
10 (44)
12 (49)
.36
.32
.29
.32
.29
.27
.34
.31
.28
.32
.28
.26
.I9
.18
.17
.15
.14
.13
-
.
-
-
STONE
8
12
16
24
ADOBE-BLOCKS
OR BRICK
POURED
CONCRETE
140 lb/w ft
80 lb/w f t
30 lb/w ft
HOLLOW
CONCRETE
BLOCKS
Sand 8
G r a v e l Agg
Cinder
Agg
Lt Wt Agg
STUCCO
HOLLOW
ON
CLAY
TILE
1 0 (25)
1
+-g-g-+
1958 ASHAE Guide
Equations: Heat Gain, Btu/hr = (Area, sq ft) X (U value) x (equivalent temp diff, Table 19)
Heat Loss, Btu/hr = (Area, sq ft) x (U value) x (outdoor temp - inside temp)
‘For addition of insulation and air spacer to above walls, refer to
Table
31, page 75.
-
CII;\I”I‘I:I:
5. 111:.\‘1‘
,\Nl)
W \‘I’I~:IC
V.\I’OII
I;I>O\zi
‘1f11<1J
S’I’I~IJ(:~I‘I_Ill~:.S
1-67
TABLE 22-TRANSMISSION COEFFICIENT U-MASONRY VENEER WALLS*
FOR SUMMER AND WINTER
Btu/(hF)
(sq ft) (deg F temp diff)
All numbers in parentheses indicate weight per rq ft. Total weight per sq ft is sum of wall and finishes.
EXTERIOR
FINISH
I
INTERIOR
BACKING
-or4” stone (50’
-orPrecast
concrete
(Sand Age’
4” 8 6”
(39’ (58’
Ad
8” 8 10”
(78’ (98’
-or-
lh ”
Board
1”
Board
Plader(7) Plaster(J)
PlasterC7)
Plaster(l)
(21
141
.26
.23
.22
.25
.21
.21
.21
.I8
.18
.I6
.I4
.I4
L
.41
.37
.30
.29
.39
..,f2
.30
(17’
.35
.32
.34
.31
.25
.23
.24
.22
.19
.I5
(32’
(43’
.30
.28
.26
.28
.29
.27
.27
.25
.23
.21
-35
.:?.!
- - - -r __, .24
.28
.23
.26
.22
.21
/
.21
.20
.21
.20
.20
.I9
.17
.I7
.I4
.I3
.46
.39
.37
.41
.35
.33
.32
.28
.27
.29
.26
.25
.29
.26
.25
.27
.25
.24
.22
.21
.20
.I7
.16
.I5
Hollow
Clay Tile
4 (16’
8 .1301
~
12 (40’
.41
.31
.26
.25
.25
.24
.20
.I9
.19
.I8
.16
.13
.46
.34
.41
.31
.32
.25
.29
.23
.29
.24
.27
.22
.22
.I9
.16
.15
Concrete
(Lt Wt Ad
80 Ib/cu ft
Sand 8. Gravel
Ad
I
(Sand &
Gravel Agg’
I
I
I
I
4 (40’
a (80’
.49
.35
.42
.31
4 (20’
8 (37’
12.(53’
.36
.29
.28
.33
.28
.26
.35
.29
.27
.32
.26
.25
.26
.22
.21
.24
.21
.20
.24
.21
.20
.23
.20
.I9
.I9
.I7
.I7
.I5
.14
.13
4 (17’
8 (32’
12 (43’
.32
.27
.25
.29
.26
.24
.30
.26
.25
.2a
.25
.23
.23
.21
.20
.22
.20
.I9
.22
.20
.19
.21
.I9
.18
.18
.17
.16
.14
.13
.13
4 (23’
8 I431
12 (631
.42
.36
.34
.38
.33
.32
.40
.35
.33
.36
.32
.30
.29
.26
.25
.26
.24
.23
.27
.24
.23
.25
.23
.22
.21
.I9
.19
.16
.15
.15
1
I
I
Hollow Clay
Tile
concrete
(Lt Wt Agg’
80 lb/w ft
(Sand 8
Gravel Agg’
- o r -
Common Brick
Equations: Heat Gain, Btu/hr
Heat
l/i ”
Lt wt
.44
.37
.35
4w concrete
Block (23’
(Sand Age)
v stone (100’
(3)
‘ii “
Sand
=hft
Lt wt
.49
.41
.3a
Brick (40) _
(Sand Ad
I
J? *
Sand
4 (23’
8 (43).
12 (63’
(Lt Wt Age’
-or-
Lt wt
be
insulating
Board
Plain or
Plastered
on Furring
(Sand
8 Gravel
Atxd
Concrete Btock
(Cinder Agg’
Precast
concrete
Sand
Aw
J/“
Gypsum or
Wood Lath
Plastered
on Furring
Metal
Lath
Plastered
on Furring
n“
Plaster
on Wall
(6’
420’
Common Brick
4” Common
Gypsum
Board
(Plaster
Board’
(2)
None
Concrete
I” ‘(’
Block -- (Cinder Agg’
(Lt Wt
4” Face
Brick (43’
THICKNESS
(inches’
and
WEIGHT
(lb per
FINISH
L OSS, Btu/hr
!
= (Area, sq ft) x (U value)
= (Area,
~q ft) x (U v&e)
1958 ASHAE Guide
x (equivalent temp diff, Table 19)
x (outdoor temp
- inside temp’
*For addition of insulation and air spacer to walls, refer to Table 31. page 75.
/’
,’
’
l-68
l’;\R~I‘
TABLE
23-TRANSMISSION
COEFFICIENT
U-LIGHT
CONSTRUCTION,
I. LO,\rJ ES~I‘I~l.\TlNG
INDUSiRlAL
WALLS*1
FOR SUMMER AND WINTER
*
Btu/(hr) (sq ft) (deg F temp diff)
All numbers in parentheses indicate weight per sq ft. Total weight per sq ft is rum of wali and finishes.
INTERIOR
WEIGHT
(lb per
rq ftl
FINISH
Insulating
Board
Fiat
Iron
III
NOna
Wood
2%”
%n
EXTERIOR
FINISH
=Yifr
(21
(3)
(2)
SHEATHING
%n
Corrugated
Transite
None
‘Vi” Ins. Board
%z” Ins. Board
II)
I21
24
Gauge
Corrugated
Iron
NO”0
‘h” Ins. Board
‘5%d’ Ins. Board
$4” W o o d
(11
3/I” W o o d
Siding
(2)
(2)
(2)
01
12)
NOW
1.16
.34
.27
.55
.26
.2l
.32
.19
.17
.26
.17
.15
.36
.21
.18
1.40
.36
.28
.46
.60
.27
.22
.33
.33
.20
.17
.22
.27
.17
.I5
.19
.3a
.21
.18
.24
.58
.37
.25
.21
.27
1
.
1958 ASHAE Guide
Equations: Heat Gain, Btu/hr = (Area, sq ft) X NJ value) X (equivalent temp diff, Table 191.
Heat loss, Btu/hr = (Area, sq ft) X iU v&e) x (outdoor temp - inside temp).
*For addition of air spacer and insulation to walls, refer to Table 31, page 75.
$Values apply when sealed with calking compound between sheets, and ot ground and roof lines. When sheets are not sealed, increase U factors by 10%.
These values may be used for roofs, heat flow up-winter; for heat flow down-summer, multiply above factors by 0.8.
TABLE
24-TRANSMISSION
COEFFICIENT
U-LIGHTWEIGHT,
PREFABRICATED
FOR SUMMER AND WINTER
CURTAIN
TYPE
.WitLLS*
.,
Btu/(hr) (sq ft) (deg F temp diff)
All numbers in parentheses indicate weight per sq ft. Total weight per sq ft is sum of wall and finishes.
METAL FACING
Core
INSULATING
CORE
MATERIAL
1
2
M E T A L F A C I N G VVlT”
%” AIR SPACE (3)
(3)
Thickness (in.)
3
Core
Thickness
4
1
2
(in.)
Glass, Wood. Cotton Fibers
Paper
Honeycomb
Paper Honeycomb with Perlite Fill, Foamglar
Fiberboard
Wood Shredded (Cemented in Preformed
Slabs)
Expanded
Vermiculite
3
5
9
I5
22
7
.2l
.39
.29
.36
.3l
.34
.12
.23
.I7
.21
.I8
.20
.08
.17
.12
.15
.13
.14
.06
.I3
.09
.I2
.10
.I1
.19
.32
.25
.29
.25
.28
.l 1
.20
.15
.I9
.16
.18
;;
.I I
.l4
.12
.13
.09
.1 1
.09
.lO
Vermiculite
20
30
or
Perlite
concrete
40
60
.44
.5l
.58
.69
.27
.32
.38
.49
.19
.24
.29
.39
.15
.19
.23
.31
.35
.39
.43
.49
.23
.27
.31
.18
.21
.25
31
.14
.17
.20
.26
Equations: Heat Gain, Btu/hr = (Area, sq ft) x (U value) X (equivalent temp diff, Table 19).
Heat Loss, Btu/hr = (Area, sq ft) x [U value) X (outdoor temp - inside temp).
*For addition of insulation and air spaces to walls, refer to Table 31. page 75.
tTotal weight per sq ft =
core density Xcore thickness
12
+ 3 Ib/sq ft
.3a
1
;;
.
CHAPTER
5 . HEAT
;\i’iI) W,\T‘I-R
V,\I’OR
FLOW
TABLE 25-TRANSMISSION
TTIRIJ
1-69
STRIJCTURES
COEFFICIENT U-FRAME WALLS AND PARTITIONS*
FOR SUMMER AND WINTER
Btu/(hr) (sq ft) (deg F temp diff)
All numbers in parentheses indicate weight per sq ft. Total weight per sq ft is sum of component materials.
INTERIOR FINISH
Metal
lath
Plastered
G
%”
EXTERIOR FINISH )
SHEATHING
Wood
Panel (2 1
I
%*
.91
.6a
.48
.42
.32
.33
.30
.25
.23
.20
.42
.37
.30
.27
.23
.31
.29
.24
4” Face Brick
V
BI (43) OR
ywood (I)
01. Asphalt
Siding (2)
None, Building Paper
S/(6” Plywood (I) o r ‘A” Gyp (2)
?h” W o o d & Bldg Paper (2)
%” insulating Board (2)
‘%” Insulating Board (3)
.73
.57
.42
.38
.30
.30
.28
.23
.22
.I9
.37
.33
.27
.2s
.21
.40.
.36
.29
.27
.22
Wood Siding (3)
OR
Wood Shingles (2)
OR =/‘I Wood
Panels (3)
None, Building Paper
S/(6” Plywood (1) or %” Gyp (21
%z“ Wood 8, Bldg Paper
1/Z” Insulating Board (2)
‘%” Insulating Board (3)
.57
.48
.36
.33
.27
.27
.25
.22
.20
.18
.33
.30
.2s
.23
.20
.35
.31
.26
.24
.21
.24
.22
.19
.I8
.16
1%;53, 1;‘;;;;;;;;;~
Inrulated,Siding
j!
(4) %z” Insulating Board (3)
Single Partition (Finish on one side only)
Double Partition (Finish on both sides)
.43
.24
Woo:L.th
Plastered
T-
Insulating
Board
Plain
0,
Plastered
I”
ocwd (2) Board (4)
None, Building Paper
S/(6” Plywood (I) o r l/S” Gyp (2)
%O” Wood 8. Bldg Paper (2)
‘h ” Insulating Board (2)
%” Insulating Board (31
&wood (I) or %” G y p (2)
Gypsum
(
OR Asbestos
Cement Siding (1)
OR Asphalt
Roll Siding (2)
%/(a”
%I”
.27
.25
.21
.29
.26
.22
.21
.I8
.20
.19
.17
.I6
.14
.33
.30
.25
.24
.20
.26
.24
.21
.20
.17
.19
.I8
.I6
.15
.I4
.30
.27
.23
.22
.I9
.24
.18
I
.22
.I9
.18
.I6
.17
.I5
.14
.I3
,
.28
.25
.22
.20
.18
.21
.I9
.17
.16
.I5
.16
.I5
.I4
.13
.I2
.60
.34
.36
.19
.23
.12
.31
.28
.24
.22
.I9
.32
.29
.24
.23
.I9
1958 ASHAE Guide
Equations: Walls-Heat Gain, Btu/hr = (Area, sq ft) X IU value) X (equivalent temp diff, roble 19).
-Heat Loss, Btu/hr = (Area, sq it) X (U value) X (outdoor temp-inside temp).
Partitions, unconditioned rpoce adjacent--Heat Gain or Loss,
Partitions, kitchen or boiler room
Btu/hr = (Area sq ft) X
odjocent-Heat Gain, Btu/hr = (Area
sq
(U value)
X (outdoor temp-inside temp-5 F).
ft) X (U value)
x (actual temp diff or outdoor &mp-inside temp + 15 F lo 25 F).
“For addition of insulation and air spacer to partitions, refer to Table 31, page 75.
TABLE
26-TRANSMISSION COEFFICIENT
U-MASONRY
PARTITIONS*
FOR SUMMER AND WINTER
Btu/(hr) (sq ft) (deg F temp diff)
All numbers in parentheses indicate weight per ~q ft. Total weight per sq ft is rum of moronry
unit and finish X 1 or 2 (finished one or both rides).
FINISH
THICKNESS
(inches)
and
Both
NO.
U’EIGHT S i d e s
of
(per
UnSides
sq f t ) f i n i s h e d F i n i s h e d
RACKING
. HOLLOW CONCRETE
BLOCK
Cinder
3 (17)
.45
One
Both
4 r2y
.40
One
Both
a d7)
.32
One
Both
1 2 (531
31
Oll.2
Both
3 (151
.3a
Olle
Agg
i
?h”
Gypsum
Board
:Plaster
Board)
w
%”
1% ”
Sand
tt wt
Sand
1t w t
A g g (6) A g g (3) Plaster(7) Plaster(3)
._79
.31
.29
30
.26
__1
.27
.29
.25
.26
.34
.31
I
.31
.29
.27
.25
11 Wf Agg
S a n d B Gravel
Am
HOLLOW
CLAY
1%”
aoord(2) Beard(4)
.24
.19
.22
.17
.22
.I7
.21
.lb
.lB
.I2
.23.lB
.21
.I6
.22
.17
.21
.15
.I7
.12
.36
.35
.34
.32
.27
/
.20
.17
j
.I7
.lb
(
.29
.2a
.27
.24
/
/
.22
.lE
.21
.I6
/
1
.21
.I6
.20
.I5
1
.27
.2b
.25
.23
1
.21
.17
.20
.15
I
.20
.lb
.19
.I5
1”
-___
.I4
.09
’
I(
.14
.09
.20
.13
.15’
.09
.19
.13
.15
.09
.17
.I2
.I4
.09
.16
.12
.13
.OB
1 2 (43) /
.25
.23
a (431
.40
One
Both
.36
.32
.39
.37
.35
.31
.2a
.21
.26
.19
.26
.19
.25
.lB
.20
.13
12 (63)
.3a
Otle
.34
.30
.3b
.35
.33
.29
.27
.21
.25
.1a
.25
.19
.24
.17
.I9
.I3
.15
.l 1
.15
.09
3 (151
.46
Otle
Both
.40
.44
.39
.31
.2a
.2B
.27
.22
.lb
4 (16)
.40
One
Both
6 (25)
.35
Olle
Both
8 (30)
.31
On.2
Both
3 (9)
.37
One
Both
.35
.32
.26
.24
.24
.23
.I9
.I5
4 (13)
.33
On-2
Both
Both
TILE
H O L L O W GYPSUA
TILE
1% ”
1% N
Sand
1t wt
Plaster/7) Plaster(l)
___~ ~
Both
i
% ”
Plaster
o n Wail
Insulating
Board
Plain or
Plastered
on Furring
%”
Gypsum or
Wood Lath
Plastered
on Furring
Metal
lath
Plastered
on Furring
.33
/
SOLID GYPSUM
PLASTER
1958 ASHAE Guide
Equations: Partitions, unconditioned space adjacent: Heat Gain or LOSS, atu/hr = (Area, Sq ft) X (U
value)
x
[outdoor
temp-inside
temp-5
FL
Partitions, kitchen or boiler room adjacent: Heat Gain or Loss, Btu/hr = (Area, sq ft) X (U value)
x (actual temp diff or outdoor temp-inside temp + I5 F to 25
*For addition of insulation and air spaces
to partitions, refer to
.
Table 31, page 75.
Fl.
(:11,\1”1’1:11
5.
TABLE
IIE.\‘I’
.\&-I) W.\‘l’l~:I~
27-TRANSMISSION
V.\I’Ol< I~l.OLV
COEFFICIENT
‘1‘11111J
U-FLAT
l-71
S’I‘RIJ(:‘I‘URES
ROOFS
COVERED
WITH
FOR HEAT FLOW DOWN-SUMMER. FOR HEAT FLOW UP-WINTER (See Equation
Btu/(hr)
All numbers in parentheses indicate weight per sq
THICKNESS
OF
DECK
(inches)
TYPE OF DECK
(sq
ft) (deg
F temp
ft. Total weight per sq
BUILT-UP
ROOFING*
of Page).
at Bottom
diff)
ft is sum of roof, finish and insulation.
INSULATION ON TOP OF DECK, INCHES
CEILING t
NO
Inrulation
(3)
1%
‘/a
(1)
2%
(2)
S u s p e n d e d Plast
(lt W t A g g on
Gypsum Board)
2 (91
None or Plaster (6)
Suspended Plaster (5)
S u s p e n d e d A c o u T i l e (2)
.27
.I8
.15
.20
.14
.12
.15
.12
.l 1
.13
.lO
.09
.l 1
.09
.08
.lO
.09
.08
.08
.08
.07
3 (13)
None or Plaster (6)
Suspended Plaster (5)
S u s p e n d e d A c o u T i l e (2)
.21
.15
.13
.16
.12
.l 1
.I3
.l 1
.\o
.l 1
.09
.OfJ
.lO
.08
.08
.09
.OE
.07
.08
.07
.06
4 (16).
None or Plaster (6)
Suspended Plaster (5)
Suspended Acou Tile(Z)
.17
.13
.12
.14
.l 1
.lO
.i 1.
.lO
.09
.lO
.08
.07
.09
.08
.07
.08
.07
.06
.07
.06
.05
2
None or Plaster (6)
Suspended Plaster (5)
S u s p e n d e d A c o u T i l e (2)
.32
.21
.17
.22
.17
.13
.17
.13
.I2
.14
.l 1
.lO
.12
.lO
.09
.lO
.09
.08
.09
.08
.07
3 (15)
None or Plaster (6)
Suspended Plaster (5)
S u s p e n d e d A c o u T i l e (2)
.27
.19
.I5
.19
.15
.12
.15
.I3
.ll
.13
.l1
.09
.l 1
.lO
.08
.lO
.09
.08
.08
.07
A (19)
None or Plaster (61
Suspended Plaster (5)
S u s p e n d e d A c o u T i l e (2)
.23
.17
.lA
.17
.13
.12
.lA
.lO
_ .12
.l 1
.12
.lO
.09
.09
.08
.09
.OE
.08
.21
.16
.13
116
.13
.ll
.13
.l 1
.lO
.ll
.09
.09
.lO
.09
.08
.09
.08
.07
N
Gypsum Slab on %”
Gypsum Board
N
Wood
(11)
I I
i I
1 (3)
None or Plaster (6)
Suspended Plaster (5)
S u s p e n d e d A c o u T i l e (2)
2 (5)
None or Plaster (6)
Suspended Plaster (5)
S u s p e n d e d A c o u T i l e (2)
3 (81
None or Plaster (61
Suspended Plaster (5)
S u s p e n d e d Acou T i l e (2)
.oa
,
.oa
.07
.07
.08
.07
.06
1958 ASHAE Guide
Equations:
Summer-(Heat
F l o w Down) Heat
Gain, Btu/hr = (Area, rq ft) X VJ v a l u e ) X ( e q u i v a l e n t t e m p d i f f , T a b l e
Winter-(Heat Flow Up) Heat Loss, Btu/hr = (Area, rq ft) X (U
*For a d d i t i o n o f a i r spaces or inrulotion
20).
value x 1.1) X (outdoor temp-inside temp).
to roofs, refer to Table 31, page 75.
tFor suspended I%” i n s u l a t i o n b o a r d , p l a i n (.6) o r w i t h I/” sand aggregate plaster
(5).
use values of suspended ocou
tile.
TABLE
28-TRANSMISSION COEFFICIENT
FOR HEAT FLOW DOWN-SUMMER. FOR HEAT FLOW
Btu/(hr) (sq ft projected area) (deg
All numbers in parentheses indicate weight per
PITCHED
U-PITCHED
F
temp diff)
sq f t . T o t a l w e i g h t p e r
sq f t i s r u m o f c o m p o n e n t m a t e r i a l s .
CEILING’
ROOFS
‘/‘a” Gypsum
Woo: Lath
Plastered
l/i ”
Sand
EXTERIOR SURFACE
ROOFS*
UP- WINTER (See Equation at Bottom of Page)
‘/a ”
tt wt
Plaster Plaster
(5)
(2)
SHEATHING
Insulating
Board Plain or
‘Ii” S a n d A g g
Plastered
Acoustical
Tile
on Furring
01
J/O Gypsum
(2)
1”
Board
(4)
l/i n
Tile
(2)
J/4”
Tile
(3)
% It
Board
Bldg paper on %/16”
Asphalt
Shingler, (2)
Asbestos-Cement
Shingles (3)
Sl
.27
30
30
.23
.26
.59
.20
.34
.29
.29
.2a
.22
.17
.23
.21
.25
.25
.24
.20
.I6
.21
.I9
.37
.33
.33
.3 I
.25
.I8
.25
.22
.32
.27
Arph:;t Roll
Roofina (1)
Eldg paper on %z”
wood sheathina (3)
.45
.25
.29
.31
.28
.20
.27
.22
.17
.22
.20
Slates (8)
T i l e (10)
B l d g p a p e r o n +‘kn
plywood (2)
.64
.29
.36
.38
.34
.35
.47
.26
.19
.26
.23
Sheet letal (1)
B l d g p a p e r o n %i
wood sheathing (3)
I------
.48
.25
.29
.31
.28
.28
.27
.22
.I7
.23
.20
Bldg paper on
1“ x 4” strips (I I
.53
.26
.31
.33
.30
.30
.2a
.23
.17
.24
.21
.41
.23
.27
.29
.26
.27
.25
.21
.16
.2l
.34
.21
.24
.25
.23
.23
.22
.19
.15
.19
Wood
Shingles (2)
B l d g paper
,
’
o n %”
plywood (2)
Bldg paper on ?/Ls
wood sheathing (31
I------
.
.19
.17
1958 ASHAE Guide
E q u a t i o n s : S u m m e r ( H e a t F l o w D o w n ) H e a t G a i n , Btu/hr = ( h o r i z o n t a l p r o j e c t e d area, sq ft) X (U v a l u e ) X ( e q u i v a l e n t t e m p d i f f , T a b l e 2 0 ) .
W i n t e r ( H e a t F l o w U p ) H e a t L o s s , Btu/hr = ( h o r i z o n t a l p r o j e c t e d a r e a , sq ft) X (U v a l u e X 1 . 1 ) X ( o u t d o o r t e m p - i n s i d e t e m p ) .
*For addition of air spaces or insulation for above roofs, refer to
Table 31,
page 75.
TABLE 29-TRANSMISSION COEFFICIENT U-CEILING AND FLOOR, (Heat Flow Up)
Based on Still Air Both Sides, Btu/(hr) (sq ft) (deg F temp diff)
All numbers in parentheses indicate weight, per sq
I-
ft. Total weight per sq
MASONRY CEILING
Suspended or Furred
Not Furred
FLOOR
Iah”
Wood Block
on Slob
1CONCRETE
( SUBFLOOR
! 4 (,y
Sand Agg 1 6 (60)
1
,;
Lt Wt Agg
80 Ib/ft”
or
2 122)
Sand Agg
on
5/a” Plywood
on
2” x 2” Sleepers
2 (16)
4 (29)
6 (42)
Floor Tile
$4” Linoleum
ACOUStiCol
THICKTile
NESS
NOW
Glued
(inches) or
,,*fl
and
%r
WEIGHT Sand
1t wt
% u
Y-4 ”
(lb per Plaster Plaster Tile
Tile
sq ff)
4 (42)
6 (62)
8 (82)
1 0 (102)
Lt Wt Agg
80
lb/f@
2 (19)
4 (31)
6 (44)
2 (24)
%“Hordwood
Sand
Agg
0”
%z”
Subfloor
on
2” x 2” Sleepers
Lt Wt
80 lb/f@
Agg
4 (44)
6 (64)
8
10
2
4
6
(84)
(104)
(20)
(33)
(46)
(51
.44
.41
p3f3)
(3)
I
Ill
1 .36
.34
3;
.29
.28
3;.
.25
.24
2;
.31
.25
.21
.28
.27
.26
.25
.24
.25
.21
.18
.23
.23
.22
.21
.20
1:;
.27
.22
.19
.26
.25
.24
.23
.22
.22
.19
.16
.24
.20
.17
.23
.22
.21
.21
.20
.20
.17
.I5
.20
.I7
.I5
.20
.19
.19
.18
.17
.17
.15
.I4
.18
.I6
.14
.18
.17
.17
.16
.16
.16
.14
.13
,
?h” Gypsum
or
Wood Lath
Plastered
Metal
Lath
Plastered
%”
=A/,”
1%”
Insulating
Board Plain or
‘/a” Sand Agg
Plastered
.31
20
2;
.28
2;
i
1;:
.25
1
.19
.31
.30
.28
.27
.26
.26
.22
.18
.25
.24
.23
.22
.21
.21
.18
.16
.18
.28
.27
.26
.25
.24
.24
.20
.17
.23
.22
.21
.21
.20
.20
.17
.I5
.16
.21
.20
.19
.19
.18
1 .32
114
.38
’
1;;
.27
.23
.I9
.32
.30
.29
.27
.26
.26
.22
.19
.25
.24
.23
.22
.22
.22
.lB
.I6
.21
Acoustical Tileon Furring
Vi” :;prum
‘vi n
Sand
l.t wt
Sand
Lt wt
‘vi *
Plaster Plaster Plaster Plaster By.;;d
(7)
(31
(51
(2)
ClJ
.36
.28
.23
.32
.31
.29
.28
.27
,
ft is rum of ceiling and floor.
1% ”
Tile
(1)
?firr
Tile
(1)
.16
.16
.22
.22
:;:,
.i
.19
:;;
.20
.17
.I5
.I8
.lB
.18
.17
.17
.17
.15
.I3
.I6
.16
.16
.I5
.15
.18
.16
.I4
.I7
.17
.16
.16
.15
.I5
.14
-12
.I5
.I5
.14
.14
.14
.14
.12
.l 1
.22
.21
1;;
.32
.26
.21
.30
.28
.27
.26
.25
.25
.21
.18
.24
.23
.22
.21
.21
.21
.18
.16
1”
Board
(41
2;
.19
.I7
.15
.18
.I8
.17
.17
.16
.16
.14
.13
.16
.16
.15
.15
.14
.14
.I3
.12
/
:;:
.I5
.I3
.I2
.14
.I4
.14
.I3
.I3
.13
.12
.l 1
.13
.13
.I2
.12
.12
.12
.I 1
.099
.15
.13
.12
FRAME CONSTRUCTION CEILING
Not Furred
Acoustical
Tile
Glued
NolIe
FLOOR
1 SUBFLOOR
‘/2 “
Tile
=Yiv
Tile
(1)
(1)
Suspended or Furred
Metal
Lath
Plastered
Yirf
Sand
vi“
1t wt
3/s” Gypsum
Woo:L.th
Plastered
vi“
Sand
‘A2 ”
Lt wt
Insulating
Board Plain or
1%” Sand Agg
Plastered
‘h ”
Plaster Plaster Plaster Plaster Board
(7)
j (3)
(21
(2)
(5)
Acoustical Tile
on Furring
or
J/s” Gypsum
1”
I/$ ff
=/qn
Board
(4)
Tile
(1)
Tile
II)
?h” Wood (2)
l/d” Hardboard on
3/g” Insulating Board
.27
.20
.21
.19
.22
.28
.20
.20
.26
.19
.20
.26
.19
.17
.38
.24
.I8
.19
.17
.19
25~” Wood (21)
.24
.I8
.20
.16
.I4
.I6
.I3
.I7
.21
.I6
.l-;
.19
.15
‘s%” Wood (5)
.33
.22
.24
.I7
.21
.I6
.25
.I8
.23
.I7
.23
.I7
.22
.17
.I8
.15
.15
.12
.19
.15
.17
.14
?h”
.28
.20
.20
.16
1 .21
.16
.20
.16
.I7
.14
.I4
.12
.I8
.14
.I6
.I3
Wood (5)
2” Wood (81
1958
E q u a t i o n s : H e a t f l o w u p , U n c o n d i t i o n e d s p a c e b e l o w : H e a t G a i n , Btu/hr = (Area, sq f t ) X (U
K i t c h e n o r b o i l e r r o o m b e l o w : Heat G a i n , Btu/hr = (Area, sq ft) X (U value)
value)
ASHAE G u i d e
x (outdoor temp - inside temp - 5 F).
x (actual temp diff, or outdoor temp - inside temp + I5 F to 25 F).
1-74
TABLE JO-TRANSMISSION COEFFICIENT U-CEILING AND FLOOR, (Heat Flow Down)
Based on Still Air Both Sides, Btu/(hr) (sq ft) (deg F temp diff)
All
numbers in parentheses indicate weight per sq ft. Total weight per sq ft is rum of ceiling and floor.
F u r r e_d _
r--~N o t-T
I
THICKNESS
:inches)
and
WEIGH1
(lb per
sq ft)
2
4
6
8
IO
Sand Agg
or
$6” linoleum
or
Floor Tile
CEILING
.- MASONRY
.___~
Suspended or Furred
-.-~~-
,-.--
Acoustical
NOlV3
or
% ”
Sand
lostet
til
.48
.44
.4l
-39
i i
.43
.40
.37
.35
.31
.30
.26
.25
32
.31
- ---1~~-‘~-T~I-nru a na
.29
.28
.30
.28
.28
.27
.23
.22
.23
.22
.22
.21
20
.20
.20
.I9
.I9
.I8
.20
.17
.I5
.18
.16
.I4
.20
.19
.18
.18
.I7
.18
.17
.17
.16
.16
(42)
.17
.15
.I4
.16
.14
.I3
(221
(42)
(62)
(82)
(102
.20
.19
.I8
.I8
.17
.I7
.I7
.I6
.16
.I5
.17
.15 .
-. .14
.17
.16
.16
.I5
.I5
.15
.I4
.I3
Tile
(1)(1) j
Tile
(1)(II
(191
(39)
(59)
(79)
(99)
.17
.I7
2 (15)
4 (28)
6 (41)
2
4
6
8
10
4cousticol
Tile
on Furring
or
?/a” Gypsum
4
c2Ol
(40)
(60)
(80)
(100
2 (16)
_-Floor Tile
or
l/8” Linoleum
on
?‘a” Plywood
on
2” x 2” Sleepers
%”
4 (29)
6
Sand Agg
I
Lt Wt Agg
8 0 Ib/ft3
r-
Hardwood
2
4
6
8
10
Sand Agg
I
-I
VIZ” ::bfloor
L--on
Lt Wt Agg
2” x 2” Sleepers
8 0 Ib/ft’
2 (19)
4 (31)
6 (441
.2l
__.21
.18
.16
.I9
.19
.I6
.I4
I
I%” Ceramic Tile
on, ‘,I/*”
rrmrnt
on1
4x” Cement
‘%6” Hard
‘=/(6”
Hardwood Floor
or L1no
or
linoleum
.--... on
-..
%” Plywood
Plvwnnd
Vi”
I,
‘%
Linoleum on
Vi” Hordboard on
‘:,
% Insulating
%”
Insulating Board
. . I . - I - - . -
;
%”
- . .
I- -.
I?!
/
.20
I
.20
.19
t
.__~
.20
.I9
.17
.15
.2l
.18
(24)
(44)
(6-i)
(84)
(104
.26
.25
.24
.23
.22
.25
.24
.23
.22
.21
.20
.20
.19
.19
.10
.18
.18
.17
.17
.I6
.20
.20
.19
.I9
.18
.I9
.10
.I8
.I7
.17
2 (20)
4 (33)
6 (46)
.22
.19
.I6
.21
.I8
.l6
.18
.I6
.I4
.16
.14
.13
.18
.16
.14
.I7
.I5
.13
2
4
6
0
10
I
,
I
I
SUBFLOOR
NOW
.20
.22
.I9
(
Tile
(I)
(1)
1
Tile
(I)
(1)
.17
.13
.17
.15
.13
.13
.12
.I 1
.I6
.13
.I6
.13
.I5
.13
.15
.I2
.14
.12
--___
.14
.13
.I2
-c
FRAME CONSTRUCTION CEILING
/
Plaster Plaster Plaster Plaster Board
Board
(7)
j (3)
/ (5)
(5)
/ (2)
j (2)
1 (41
(3)
(41 j
(2)
(2)
None
.21
.I5
.I2
Wood (2)
72” Wood
.12
.I 1
.10
.15
.15
.14
.14
.14
.15.14
.13
.12
.12
.ll
Suspended or Furred
Not Furred
FLOOR
.21
.2l
.18
.I6
.3l
.20
.15
.I5
.12
.12
.I 1
(24)
‘K” Wood
- - - (5)
Y Wood
Wnnrl
2”
(7)
--. -. ~-p----t
%z” Wood (5)
4::
$#j;,;- (8)
2” Wood
.14
.I2
.I2
.I4
.13
.I0
.12
.I I
---~___.I I
.10
.14
.1 1
1958
ASHAE
Equations Heot flow down, unconditioned space above: Heat Gain, Btu/hr = (Area. 54 ft) X (U value) X ( outdoor temp - inside temp - 5 FL
Kitchen above: Heat Gain, Btu/hr = (Area, sq ft) X (U v&e) X (actual temp diff, or outdoor temp - inside temp + 15 F to 25 FL
.
.27
.I7
.14
.13
.1 1
Guide
’
l-75
CH/\I”I‘ER 5. HEAT AND WATER V.\I’OK FLOW THRU STRUCTURES
TABLE Jl--TRANSMISSION
COEFFICIENT U-WITH INSULATION & AIR SPACES
SUMMER AND WINTER
Btu/(hr) (sq ft) (deg F temp
diff)
Addition of Reflective Sheets to Air Space (Aluminum Foil Average Emisrivity
u Value
Before
Adding
Intul.
Wall,
Ceiling,
Roof
Floor
-
Direction
Addition of
Fibrous Insulation
Thickness
Winter and Summer
Horizontal
4dded
to one
>r both
sides
(Inches)
of
Added
to one
or both
sides
Olle
sheet
in air
Space
TWO
sheets
in air
lp.Xe
Added
to one
or both
sides
One
sheet
in air
IpCXe
TWO
sheets
in air
SPClC.3
.05
.05
.OS
.05
.05
.05
35
.36
35
.34
.33
.32
.20
.20
.20
.I9
.19
.I9
.I4
.I4
.14
.I4
.14
.I3
.04
.04
.04
.04
.04
.31
.30
.29
.28
.27
.18
.18
.I8
.17
.17
.I3
.I3
.13
.13
.I2
.04
.04
.04
.04
.04
.26
.25
.24
.23
.22
.17
.16
.16
.I5
.15
.I2
.I2
.12
.11
.l 1
.04
.04
.04
.04
.04
.20
.19
.18
.16
.15
.I4
.13
.I3
.12
.l I
.I0
.I0
.lO
.09
.09
.04
.04
.04
.03
.03
.14
.13
.I2
.I0
.09
.l 1
.lO
.09
.08
.07
.08
.08
.07
.07
.06
2
3
.60
.sa
.56
.s4
.52
30
.08
.08
.08
.08
.08
.08
.38
.37
.36
.36
.35
.34
.34
.33
.32
.3l
.30
.29
.I8
.18
.I8
.17
.I7
.17
.l 1
.l 1
.l 1
.l 1
.lO
.I0
.12
.I2
.l 1
.I I
.I I
.l 1
.06
.06
.06
.06
.06
.06
.48
.46
.44
42
40
.I7
.17
.I7
.I6
.16
.I I
.lO
.lO
.lO
.lO
08
.08
.07
.07
.07
.33
.32
.31
.30
.29
.28
.28
.27
.26
.26
.I6
.16
.16
.I5
.I5
.I0
.lO
.I0
.lO
.lO
.l I
.I I
.I 1
.I 1
.I0
.06
.06
.06
.06
.06
.3a
.36
.34
.32
.30
.I6
.15
.15
.15
.I4
.I0
.lO
.I0
.lO
.09
.07
.07
.07
.07
.07
.28
.27
.26
.25
.23
.lO
.I0
.I0
.lO
.lO
.06
.06
.06
.05
.05
.28
.26
.24
.22
.20
.14
.I3
.13
.12
.12
.09
.09’
.09
.08
.08
.07
.07
.07
.06
.b6
.22
.21
.20
.18
.17
.09
.09
.09
.08
.08
.05
.05
.05
.05
.05
.lB
.16
.l 1
.lO
.09
.08
.07
.08
.07
.07
.06
.06
.06
.06
.05
.05
.05
.15
.I4
.12
.I I
.09
.08
.07
.07
.06
.06
.05
.05
.04
.04
.04
.14
.I2
.lO
Insulation
Added
Air
Space
Added
AIR SPACES
INSULATION
Dl<lOER
.lO
.09
.08
.08
.07
Reflective Sheets
Added to One or
Both Sides
AIR SPACE
REFLECTIVE
SHEETS
Winter
UP
TWO
sheets
in air
*pWX
.l 1
.l I
.l 1
.ll
.l 1
.I 1
.14
.12
.l 1
.I0
.08
= .05)
Flow
DOWll
.I9
.19
.18
.I8
.18
.18
-
Heat
.07
.07
.06
.06
.05
:
I
Reflective Sheet
in
Air Space
AIR SPACES
REFLECTIVE
SHEETS
‘Checked for summer conditions for up, down ond horizontal heat flow. Error from above values is less than 1 yo.
1958 ASHAE
Reflective Sheets
in
Air Space
AIR SPACES
REFLECTIVE
SHEETS
,
Guide
PART I. LOAD ESTIMATING
l-76
TABLE 32-TRANSMISSIONCOEFFICIENT U-FLAT ROOFS WITH ROOF-DECK INSULATION
SUMMER
AND
WINTER
Btu/(hr) (sq ft) (deg F temp diff)
Addition
U VALUE OF ROOF
BEFORE ADDING
ROOF DECK
INSULATION
1%
350
.50
.40
33
.29
.26
.22
.21
.I9
24
_-L- _ --.2 1
.I9
.16
.12
.09
&J
\
~&p-.25
-20
.15
.lO
of Roof-Deck
Thickness (in.)
.___.--.
1
1%
Insulation
__3
2
2%
.17
.I6
.15
.I4
.14
.13
.I2
.12
.l 1
.lO
.I0
.09
.18
.16
.15
.14
.13
.12
.I2
.I2
.l 1
.I0
.10
.09
.09
.09
.08
.13
.I 1
.08
.I 1
.09
.07
.I0
.08
.07
.09
.08
.06
.08
.07
.05
TABLE 33-TRANSMISSIONCOEFFICIENT U-WINDOWS, SKYLIGHTS,
DOORS & GLASS BLOCK WALLS
.
Btu/(hr) (sq ft) (deg F temp diff)
GLASS
Vertical
Single
Air Space Thickness (in.)
Horizontal
Triple
Double
‘A
,WWithout
Sto
m
Windows
1.131‘
W i t h S t o r m - c&. mdows lm
Glass
‘h
Oh’
%-4
0.55
‘A
0.53
1%
0.4 1
0.36
Single
%-4
0.34
Summer
0.86
0.43
Glass
Double (%“)
Winter
Summer
Winter
1.40
0.64
0.50
0.70
DOORS
Nominal
Thickness
of Wood (inches)
u
Exposed Door
U
With Storm Door
1
1%
1%
1%
0.69
0.59
0.52
0.51
0.35
0.32
0.30
0.30
2
0.46
0.38
0.33
1.05
0.28
0.25
0.23
0.43
2’/2
3
Glare (%”
Herculite)
HOLLOW
GLASS
BLOCK
WALLS
Description*
U
5%x5%x3%” Thick-Nominal Size 6x6x4 (14)
7%/7%x3%” Thick-Nominal Size 8x8~4 (14)
11 %x1 1 %x3%” Thick-Nominal Size 12x12~4 (16)
7%x7%x3%” Thick with gloss fiber screen dividing the cavity (14) .
11 %x1 1 %x3%” Thick with glass fiber screen dividing the cavity (16)
-
0.60
0.56
0.52
0 . 4 8
0.44
1958 ASHAE
Equation: Heat Gain or Loss,
Btu/hr = (Area, rq ft) X (U
*Italicized numbers in parentheses indicate weight in lb per
.
value) x (outdoor temp
sq ft.
,,_
- inside temp)
Guide
CHAI’TEK
,5. HEA~I‘
.\NI)
CV.\~I‘I<K
V,\I’OK
I:I.OW .I‘HKIJ
CALCULATION OF TRANSMISSION
COEFFICIENT U
For types of construction not listed in Tabtes 21
thru 33, calculate the U value as follows:
1. Determine the resistance of each component of
a given structure and also the inside and outdoor air surface films from Table 34.
2. Add these resistances together,
l-77
S’I‘I~IJ~:‘I’IJI~I‘S
Example 6 - Calculation
of
U Value
Given:
A wall as per Fig. 27
R = rl + yI + T,~ + . . . . . r,
FIG. 27
3. Take the reciprocal, U =k
Basis of Table 34
- Thermal Resistance R, Building and Insulating Materials
Find:
Transmission
Table 34 was extracted from the 1958 ASHAE
Guide and the column “weight per sq ft” added.
Solution:
Refer to Table 34.
&o of Table 34
lermql Resistance R, Building and Insulating Materials
The thermal resistances for building materials are
listed in two columns. One column lists the thermal
resistance per inch thickness, based on conductivity,
while the other column lists the thermal resistance
for a given thickness or construction, based on conductance.
I.
2.
3.
4.
5.
coefficient
- OUTDOOR W ALL
in
summer.
Construction
Outdoor air surface (7% mph wind)
Stone facing, 2 in. (2 X .OS)
Hollow clay tile, 8”
Sand aggregate plaster, 2 in. (2 X .20)
Inside air surface (still air)
-..
Total
Resistance
U = + = & = 0.30 Rtu/(hr)(sq
Resistance
R
0.25
0.16
1.85
0.40
0.68
ft)(deg F)
I’;\R’I‘
l-78
TABLE
34-THERMAL RESISTANCES
R-BUILDING
AND
I .
INSULATING
LOAD ESTIM,\TING
MATERIALS
(deg F per Btu) / (hr) (sq ft)
RESISTANCE R
THICKNESS
(in.)
DESCRIPTION
MATERIAL
DENSITY
(lb per
C” f0
WEIGHT
(lb per
l I ft)
Per Inch
Thickness
1
-iF
For Listed
Thickness
1
e
BUILDING MATERIALS
BUILDING
BOARD
Boards,
Panels,
Sheathing,
etc
Asbestos-Cement Board
Asbestos-Cement Board
Gypsum or Plaster Board
Gypsum or Plaster Board
Plywood
Plywood
Plywood
Plywood
Plywood or Wood Panels
W o o d F i b e r B o a r d , Lominoted
Wood
Wood
Wood,
Wood,
Fiber,
Fiber,
Fir or
Fir or
1%
%
‘55
%
%
%
%
or H o m o g e n e o u s
Hardboard Type
Hardboard Type
Pine Sheathing
Pine
‘/I
?YJl
1%
-
-
0.25
-
-
120
120
50
50
34
34
34
34
34
26
31
1.25
1.58
2.08
0.71
1.06
1.42
2.13
-
1.25
2.38
2.00
0.03
0.32
0.45
0.31
0.47
0.63
0.94
-
65
65
32
32
1.35
2.08
4.34
0.72
-
0.18
0.98
2.03
-
-
BUILDING
PAPER
Vapor Permeable Felt
Vapor Seal, 2 Layers of Mopped 15 lb felt
Vapor Seal, Plastic Film
-
-
-
0.06
0.12
Negl
WOODS
Maple, Oak, and Similar Hardwoods
Fir. Pine. and Similar Softwoods
4.5
32
-
0.9 1
1.25
-
MASONRY
UNITS
Brick, Common
Brick, Face
Clay Tile, Hollow:
I Cell Deep
1 Cell Deep
2 Cells Deep
2 Cells Deep
2 Cells Deep
3 Cells Deep
Concrete Blecks, Three Oval Core
Sand 8. Gravel Aggregate
Cinder
Aggregate
Lightweight Aggregate
( E x p a n d e d S h a l e , C l a y , S l a t e or
Slag; Pumice)
Gypsum Partition Tile:
3”xl2”x30” solid
3”xl2”x30” 4-cell
4”xl2”x30” 3-cell
L
Stone, Lime or Sand
.
4
4
120
130
40
43
3
4
6
a
10
I2
60
40
50
45
42
40
15
16
25
30
35
40
3
4
6
8
12
76
69
64
64
63
19
23
32
43
63
3
4
6
8
12
68
60
54
56
53
17
20
27
37
53
3
4
8
12
60
52
48
43
15
17
32
43
-
3
3
4
45
35
30
11
9
13
-
1.26
1.35
1.67
0.08
-
150
-
-
30
.44
0.80
1.1 ?
1.52
1.85
2.22
2.50
0.40
0.71
0.91
1.11
1.28
0.86
1.1 I
1 so
1.72
1.89
1.27
1.50
2.00
2.27
.
(:M.\I”I‘l~li
5. HL\‘I’
TABLE
.\NI) LV.\
I I;li V.\I’OI<
I:I.OW
34-THERMAL RESISTANCES
‘I’IlllIJ
l-79
S’I‘IIIJ(:‘I’tIIII~S
R-BUILDING
AND
INSULATING
MATERIALS
(Contd)
(deg F per Btu) / (hr) (sq ft)
RESISTANCE
MATERIAL
THICKNESS
(in.)
DESCRIPTION
DENSITY
(lb p e r
C” ft)
WEIGHT
(lb p e r
sq ft)
Per Inch
Thickness
1
-ii-
BUILDING
MASONRY
MATERIALS
Concretes
PL .TERIN6
MATERIALS
ROOFING
SIDING
MATERIALS
(On Flat Surface)
tement Mortar
Gypsum-Fiber Concrete
12’/a’$& w o o d c h i p s
MATERIALS ,
((:ONT.)
116
Bi’Yx%
gypsum,
51
-
.ightweight Aggregates
Including Expanded
Shale, Clay or Slate
Expanded Slag; Cinders
Pumice; Perlite; Vermiculite
Also, Cellular Concretes
120
100
80
60
40
30
20
S a n d & Gravel or Stone Aggregate (Oven Dried)
S a n d & Gravel or S t o n e A g g r e g a t e ( N o t D r i e d )
stucco
140
140
116
Cement Plaster, Sand Aggregate
Sand
Aggregate
Sand
Aggregate
116
116
116
4.8
7.2
Gypsum Plaster:
lightweight Aggregate
Lightweight Aggregate
lightweight Aggregate on Metal Lath
Perlite Aggregate
Sand
Aggregate
Sand
Aggregate
Sand
Aggregate
Sand Aggregate on Metal Lath
Sand Aggregate on W o o d L a t h
Vermiculite Aggregate
45
45
45
45
105
105
105
105
105
45
~.BB
2.34
2.80
4.4
5.5
6.6
-
Asbestos-Cement Shingles
Asphalt Roll Roofing
Asphalt Shingles
Built-up
Roofing
Slate
Sheet Metal
Wood Shingles
120
70
70
70
201
40
-
Shingles
W o o d , 16”, 7%” exposure
W o o d , D o u b l e , 16”. 12” exposure
W o o d , P l u s lnsul Backer Board, ?&”
Siding
A s b e s t o s - C e m e n t , ‘A” lapped
Asphalt Roll Siding
A s p h a l t lnrul S i d i n g ‘An Board
Wood Drop 1”xB”
Wood: Bevel ‘/z”xB”;,lapped
W o o d I Bevel, Y/x10 , lapped
W o o d , P l y w o o d , 3/s”, lapped
%
‘Ii
Asphalt Tile
Carpet and Fibrous Pad
Carpet and Rubber Pad
Ceramic Tile
Cork Tile
Cork Tile
Felt, Flooring
Floor Tile
Linoleum
Plywood Subfloor
Rubber or Plastic Tile
TfXraZIO
Wood Subfloor
Wood, Hardwood Finish
2.2
8.4
-
-
-
-
-
-
Structural Glass
FLOORING
MATERIALS
-
120
25
25
-
1.25
0.26
-
80
34
110
140
32
45
0.83
1.77
1.15
11.7
2.08
2.81
R
For Listed
Thicknert
1
c
c-
-
0.20
0.60
0.19
0.28
0.40
0.59
0.86
1.11
1.43
0.11
0.08
0.20
0.20
-
0.67
0.1B
0.59
-
1
1-
610
&5
0.32
0.39
0.47
0.09
0.11
0.13
0.40
-
Negl
-
0.21
0.15
0.44
0.33
0.05
0.94
-
0.87
1.19
1.40
2.22
-
0.21
0.15
1.45
0.79
0.81
1.05
0.59
0.10
0.04
2.08
1.23
0.08
0.28
0.06
0.05
0.08
0.78
0.02
0.08
0.98
0.68
I
l'/\R'I I. LoAD I:s'I‘IM.\'I‘INC;
l-80
TABLE
34-THERMAL
RESISTANCES
R-BUILDING
AND
INSULATING
MATERIALS
(Contd)
(deg F per Btu) / (hr) (sq ft)
-I
I
I-
RESISTANCE
MATERIAL
THICKNESS
(in.)
DESCRIPTION
DENSITY
(lb p e r
cu
ft)
/
WEIGHT
(lb per
‘ 4 ffl
Per Inch
Thickness
1
k
I
INSULATING MATERIALS
BLANKET
AND
BOARD AND
BATT*
SLABS
LOOSE FILL
ROOF INSULATION
I
L
-7
AIR SPACES
AIR FILM
Still Air
Cotton Fiber
I.8 - 2 . 0
3.85
Mineral Wool, Fibrous Form
Processed From Rock, Slag, or Gloss
1.5 _ 4.0
3.70
4.00
Wood Fiber
Wood Fiber, Multi-layer Stitched Expanded
3.70
Glass Fiber
4.00
Wood or Cane Fiber
Acoustical Tile
Acoustical Tile
Interior Finish (Tile, Lath, Plank)
Interior Finish (Tile, Lath, Plonk)
22.4
22.4
15.0
Roof Deck Slab
Sheathing (Impreg or Coated)
Sheathing (Impreg or Coated)
Sheathing (Impreg or Coated)
%
“h
0.62
20.0
20.0
20.0
0.83
9.0
6.5 _ 8.0
8.5
Macerated Paper or Pulp Products
Wood Fiber: Redwood, Hemlock, or Fir
Mineral Wool (Glass, Slag, or Rock)
Sawdust or Shavings
Vermiculite (Expanded)
2.5
2.0
2.0
8.0
1 .b2
22.0
_ 3.5
- 3.5
- 5.0
_ 15.0
7.0
All Types
Preformed, for use above deck
Approximately
Approximately
Approximately
Approximately
Approximately
Approximatley
AIR
POSITION
Horizontal
Horizontal
Horizontal
Horizontal
Horizontal
Horizontal
Horizontal
Horizontal
Horizontal
Sloping 45’
Sloping 45’
Vertical
Vertical
HEAT FLOW
Up (Winter)
Up (Summer)
Down (Winter)
Down (Winter)
Down (Winter)
Down (Winter)
Down (Summer)
Down (Summer)
Down (Summer)
Up (Winter)
Down (Summer)
Horiz. (Winter)
Horiz. (Summer)
POSITION
Horizontal
Sloping 45’
Vertical
Sloping 45”
Horizontal
HEAT FLOW
UP
UP
Horizontal
Down
CiOW”
15 Mph Wind
Any Position (For Winter)
Any Direction
7% M p h W i n d
Any Position (For Summer)
Any Direction
TI
i
‘Ii
11:
2
2%
-
15.0
Cellular Gloss
Cork Board (Without Added Binder)
Hog Hair (With Asphalt Binder)
Plastic (Foamed)
Wood Shredded (Cemented in Preformed Slabs)
2.86
-
-
2.63
-
I .3 I
2.50
3.70
3.00
3.45
I .82
-
3.57
3.33
3.33
2.22
2.08
3
3A . 4
74 - 4
9.4
1%
4
8
=/4
1 Y2
-4i
=/4 - 4
=/4 - 4
=/4 . 4
=/4 - 4
-
-
-
-
-
-
-
-
.7
-
15.6
15.6
15.6
15.6
15.6
15.6
1.3
1.9
2.6
3.2
3.9
-
-
R
For listed
Thickness
1
c
~1.19
1.78
-
*
1.32
2.06
l
-
1.39
2.78
4.17
5.26
6.67
8.33
0.85
0.78
1.02
1.15
I .23
1.25
0.85
0.93
0.99
0.90
0.89
I
0.97
0.86
0.61
0.62
0.68
0.76
0.92
_
0.17
k
0.25
‘Includes paper backing and facing if any. In corer where the insulation forms a boundary (highly reflective) of an air space, refer to Table 31, page 75
:/
CHAI’TEK
5
.
HEA’I. AND W,\~I‘tCK
V/\I’OK
I;I.OW
HEAT LOSS THRU BASEMENT WALLS AND
FLOORS BELOW THE GROUND LEVEL
The fess through the floor is normally small and
relatively constant year round because the ground
temperature under the floor varies only a little
throughout the year. The ground is a very good heat
sink and can absorb or lose a large amount of heat
without an appreciable change in temperature at
about the 8 Et level. Above the 8 ft level, the ground
temperature varies with the outdoor temperature,
with the greatest variation at the surface and a decreasing variation down to the 8 ft depth. The heat
loss thru a basement wall may be appreciable and it
is difficult to calculate because the ground temperature varies with depth. Tables 35 thru 37 have been
empirically calculated to simplify the evaluation of
heat loss thru basement walls and Hoors.
‘he heat loss thru a slab floor is large around
the perimeter and small in the center. This is because the ground temperature around the perimeter
varies with the outdoor temperature, whereas the
ground temperature in the middle remains relatively
constant, as with basement floors.
Basis of Tables 35 thru 37
-Heat Loss thru Masonry
-1‘HKU
1-81
SI‘KUCI‘UK1SS
Example 7 - Heat Loss in CI Basement
Given:
Basement - 100’ X 40’ X 9
Basement temp- 65 F tll), heated continuously
Outdoor temp - 0” F tlb
Grade line -G ft above basement floor
Walls and floors - 12 in. concrete (80 lb/cu
Solution:
I. Heat loss above ground
H = UA, (ta - to,)
= 0.18 x (ZOO + 80) X 3 X (65 - 0) = 9828 Btu/hr
2.
Heat loss
ground.
thru
walls
and
and
Walls
in
Use of Tables 35 thru 37
- Heat Loss thru Masonry Floors and Walls in Ground
Yhe transmission coefficients listed in Table 35
.may be used for any thickness of uninsulated masonry floors where there is good contact between the
floor and the ground.
The perimeter factors listed in Table 36 are used
for estimating heat loss thru basement walls and the
outside strip of basement Hoors. This factor can be
used only when the space is heated continuously. If
there is only occasional heating, calculate the heat
10~s using the wall or floor transmission coefficients
as listed in Tnbles 21 tllru 33 and the temperature
difference between the basement and outdoor air or
ground as listed in Table 37,
The heat loss in a basement is determined by adding the heat transferred thru the floor, the walls and
the outside strip of the Hoor and the portion of the
wall above the ground level.
strip
of
floor
below
= 19,100 Btu/hr
3. Heat loss thru floor
H = UAi (th - to)
= 0.05 x (100 X 40) X (65 - 55)
Total Heat Loss
where
Ground
Tables 35 thru 37 are based on empirical data.
The perimeter factors listed in Table ?6 were developed by calculating the heat transmitted for each
foot of wall to an 8 ft depth. The ground was
assumed to decrease the transmission coefficient, thus
adding resistance between the wall and the outdoor
air. The transmission coefficients were then added to
arrive at the perimeter factors.
outside
H = Lp 4 (b - toa)
= (200 + 80) X 1.05 X (65 - 0)
= 2 0 0 0 Btu/hr
= 30,928 Btu/hr
U
= Heat transmission coefficient of wall above
ground ( T a b l e 21) and floor ( T a b l e 3 5 ) i n
Btu/(hr) (sq ft) (deg F)
I
A, = Area of wall above ground, sq ft
A, = Entire floor area,
Floors
ft)
Find:
Heat loss from basement
sq
ft
Lp = Perimeter of wall, ft
Q = Perimeter factor (Table 36)
tb = Basement dry-bulb temp, F
tg = Ground temp, F, (Table 37)
t oa = Outdoor design dry-bulb temp, F
TABLE 35-TRANSMISSION COEFFICIENT UMASONRY FLOORS AND WALLS IN GROUND
(Use only in conjunction with Table 36)
Portion of Wall
exceeding B feet
below ground level
.08
*Some additional floor loss is included in perimeter factor, see Table 36.
Equations:
H e a t l o s s t h r o u g h f l o o r , Btu/hr = (area of floor, sq
X (U value) x (basement - ground temp).
ftt)
H e a t l o s s t h r o u g h w a l l b e l o w 8 f o o t l i n e , Btu/hr
= (area of wall below 8 ft line, sq ft) X (U value)
X (basement - ground temp).
NOTE: The factors in Tables 35 and 36 may be used for ony thickness
of uninsulated masonry wall or floor, but there must be a good contact
(no air space which may connect to the outdoors) between the ground
and the floor or wall. Where the ground is dry and sandy, or where
there is cinder fill along wall or where the wall has a low heat transmission coefficient, the perimeter factor may be reduced slightly.
l-82
I’AKT I. LOAD ESTIMATING
TABLE 36-PERIMETER
TRANSMISSION COEFFICIENTS PIPES IN WATER OR BRINE
FACTORS
FOR ESTIMATING HEAT LOSS THROUGH BASEMENT WALLS
AND OUTSIDE STRIP OF BASEMENT FLOOR
Heat transmission coefftcients
for copper and steel
pipes are listed in Tables 38 and 39. These coeflicicnts may be ~~scful in applications such as cold
water or brine storage systems and ice skating rinks.
(Use only in conjunction with Table 35)
Distance of Floor
From Ground Level
Perimeter
(91
2 Feet above
Factor
Basis of Tables 38 and 39
-Transmission Coefficients, Pipes in Water or Brine
.90
.60
.75
.90
At ground level
2
below
4 Feet below
6 Feet below
8 Feet below
Feet
Table 35 is for ice coated pipes in water, based on
a heat transfer film coefficient, inside the pipe, of
1 5 0 Btu/(hr)(sq Itinternal pipe surface)(deg F).
Table 39 is Sor pipes in water or brine based on a
heat transfer of 18 Btu/(hr)(sq Et external pipe surface) (deg F) in water, 14 Btu in brine. It is also based
on a low rate of circulation on the outside of the
pipe and 10 F to 15 F temperature difference between water or brine and refrigerant. High rates of
circulation will increase the heat transfer rate. For
special problems, consult heat transfer reference
books.
1.05
I .20
Equation:
Heat loss about perimeter, Btu/hr = (perimeter of wall, ft)
x (perimeter factor) X (basement - outdoor temp).
TABLE
IR
37-GROUND
TEMPERATURES
ESTIMATING HEAT LOSS THROUGH BASEMENT FLOORS
Outdoor Design Temp (F) - 30
Ground Temp (F)
1 40
- 20
- 10
/ 45
1 50
0
1 55
TABLE 3B-TRANSMISSION
+10
+20
1 60
/ 65
COEFFICIENT U-ICE COATED PIPES IN WATER
Btu/(hr) (lineal ft pipe) (deg F between 32 F db and refrig
temp)
.
Inside film coefficient = 150 Btu/(hr) (sq
Copper
Pipe
Size
(Inches
O.D.1
%
%
%
1%
Copper Pipe With
Ice Thickness (Inches)
‘vi
6.1
7.1
8.0
9.8
I
’
1%
2
4.5
5.1
5.7
6.7
3.8
3.4
4.2
4.7
5.4
if
417
it) (deg F)
Steel Pipe
Size
Nominal
(Inches)
Steel Pipe With
Ice Thickness (Inches)
%
1%
%
7.2
8.7
1
1%
10.6
1
2:
7.2
8.6
13.0
1%
2
3
4.4
5.1
5.8
3.9
4.5
5.1
6.8
5.9
3.4
3.8
4.2
4.8
TABLE 39-TRANSMISSION COEFFICIENT U-PIPES IMMERSED IN WATER OR BRINE
Btu/(hr) (lineal ft pipe) (deg F between 32 F db and refrig
Outside water film coefficient= 1 B Btu/(hr) lsq ftt)
Outside brine film coefficient= 14 Btu/(hr) (sq
Water refrigerant temp=
10 F to 15 F
(deg F)
ft) (deg F)
temp)
.
CH;\I”I‘lCK
5
.
HE.\‘I’
.\NI) W\~l‘lili V.\I’OK
I;l.OW
‘I’HKU
WATER VAPOR FLOW THRU BUILDING
STRUCTURES
Water vapor [lows thru building structures; resulting in a latent load whenever a vapor pressure
difference exists across a structure. The latent load
from this source is usually insignificant in comfort
applications and need be considcrcd only in low or
high tlewpoint applications.
Water vapor flows from high to lower vapor pressure at a rate determined by the permeability of
the structure. This process is quite similar to heat
flow, except that there is transfer of mass with water
vapor flow. As heat flow can be reduced by adding
insulation, vapor How can be reduced by vapor barriers. The vapor barrier may be paint (aluminum
or asphalt), aluminum foil or galvanized iron. It
sF--lid always be placed on the side of a structure
h
ng the higher vapor pressure, to prevent the
water vapor from flowing up to the barrier and condensing within the wall.
Basis of Table 40
-Water Vapor Transmission
thru
Various
Materials
The values for walls, floors, ceilings and partitions
have been estimated from the source references listed
in the bibliography. The resistance of a homogeneous material to water vapor transmission has been
assumed to be directly proportional to the thickness,
and it also has been assumed that there is no surface
resistance to wa.ter vapor flow. The values for, permeability of miscellaneous materials are based on
test results.
NOTE: Some of the values for walls, roofs, etc.,
l-83
S’I’I~IJ(:‘I‘IJl~I:S
have been increased by a safety factor because conclusive data is not available.
Use of Table 40
-Water Vapor Transmission
thru
Various
Materials
T&le 40 is used to determine latent heat gain
from water vapor tiransmission thru building structures in the high and low dewpoint applications
where the air moisture content must be maintained.
Example 8 - Wafer Vapor Transmission
Given:
iZ 40 ft X 40 ft X 8 ft Inhoratory on secontl floor requiring
insicle design contlitions of 40 F tll), 50y0 rh, with the olltdoor design contlitions ;It 99 I: tll). 75 F WI,. The outdoor
wall is 12 inch Ijrick with no wintlows. The partitions are
metal lath and plaster on both sides of studs. Floor and
ceiling are 4 inch concrete.
Find:
The latent heat gain from the water vapor transmission.
Solution:
Gr/lh at 95 F db, 75 F WI, = 99 (psych chart)
Gr/lh at 40 F db, 5O70 rh = 18 (psych chart)
Moisture content difference =
81 gr/lb
Assume that the dewpoint in the areas surrounding the laboratory is uniform and equal to the outdoor dewpoint.
Latent heat gain:
40 x 8
Outdoor wall = 100 X 81 X .04 (Table 40)
= 10.4 Btu/hr
40 x 40
Floor and ceilings = 2 X - X 81 x .lO
100
= 259 Btu/hr
40 X 8
Partitions = 3 X - x 81 x 1.0
100
= 777 Btu/hr
Total
Latent
Heat
Gain
= 1046.4 Btu/hr
,
I’Alil
L
1-84
TABLE 40-WATER VAPOR TRANSMISSION
I. LOi\D ESTIM;\‘1‘1N<;
THRU VARIOUS MATERIALS
PERMEANCE
Btu/(hr) ( 1 0 0 sq ft)
(gr/lb .diffl latent heat
With 2 Coots
Vapor-seal
Paint on
Smooth
Inside
Surface*
With
Aluminum
Foil Mounted
on One Side
of Poper
Cemented
to Wallt
.I2
.06
.04
.49
,075
.046
.033
-
.024
.020
.017
-
.067
.034
.40
,050
.029
-
.02 1
.016
-
Frame-with plaster interior finish
-some with asphalt coated insulating board lath
.79
.42
.I6
.I4
.029
.028
Tile-hollow cloy (face, glazed)-4 inches
-hollow clay (common)-4 inches
-hollow clay, 4 inch face and 4 inch common
.013
.24
.012
,012
.ll
.Ol I
.0091
.025
.0086
.lO
.051
2.0
SO
.40
.067
.040
.18
.14
.I3
.023
.019
.030
.028
.028
4.0
1 .o
.I9
.17
.030
.029
.02
.02
1.5
.02
.02
.018
.018
.18
.018
.018
.012
.012
.29
.012
.012
.17
.027
No Vapor
Seal Unless
Noted Under
Description
DESCRIPTION OF MATERIAL OR CONSTRUCTION
WALLS
Brick-
4
inches
- 8 inches
- 1 2 inches
-per inch of thickness
Concrete-
6 inches
- 12 inches
-per inch of thickness
I
CEILINGS
AND
FLOORS
Concrete-4
inches
-8 inches
Plaster on wood or metal lath on joist-no flooring
Plaster on wood or metal lath on joist-flooring
Plaster on wood or metal lath on joists-double flooring
PARTITIONS
Insulating Board % inch on both sides of studding
Wood or metal lath and plaster on both sides of studding
ROOFS
Concrete-2 inches, plus 3 layer felt roofing
-6 inches, plus 3 layer felt roofing
Shingler, sheathing, rafters-plus plaster on wood
Wood-l inch, plus 3 layer felt roofing
-2 inches, plus 3 layer felt roofing
or
metal
lath
MISCELLANEOUS
Air Space, still air 3% inch
1 inch
Building
Materials
Masonite-l thickness,?%
inch
-5 thicknesses
Plaster on wood lath
-plus 2 coats aluminum point
Plaster on gypsum lath
-ditto plus primer and 2 coats lead and oil paint
Plywood--% inch Douglas fir (3 ply)
-ditto plus 2 coats asphalt paint
-ditto plus 2 coats aluminum paint
--% inch Douglas fir (5 ply)
-ditto plus 2 coats asphalt paint
-ditto plus 2 coats aluminum paint
Wood-Pine .508 inch
-ditto plus 2 coats aluminum paint
-spruce, .508 inch
Insulating
Materials
Corkboard, 1 inch thick
Interior finish insulating board, l/i“
-ditto plus 2 coats water emulsion paint
-ditto plus 2 coats varnish base paint
-ditto plus 2 coots lead and oil paint
-ditto plus wall linoleum
3.6
13.0
1.1
.32
1.1
1.95
.63
-
.12
.13
.087
.I3
.27
--
.04 1
.12
.33
.20
.046
.63
5.0 3.0 .1 .17
.03 -
7.0
4.0
1.0
.06
<:FI.\1”1‘1~1<
i. ill~.\‘l‘
.\NI)
LV.\‘1‘1-I<
V.\I’OI<
I~I.OCV
1’111111
l-85
s’I’l1Il(:‘l’lJl<I:s
TABLE 40-WATER VAPOR TRANSMISSION THRU VARIOUS MATERIALS (Contd)
PERMEANCE
Btu/(hr) (100 sq ft)
(gr/lb diff) latent heat
DESCRIPTION OF MATERIAL OR CONSTRUCTION
No Vapor
Seal Uliless
Noted Under
Description
,
With 2 Coats
Vapor-real
Paint on
Smooth
Inside
Surfate*
With
Aluminum
Foil Mounted
on One Side
of Paper
Cemented
to wallt
MISCELLANEOUS
Insulating Materials. cont.
Insulating board lath
-ditto plus %” plaster
-ditto plus 1/2” plaster, sealer, rind flat coat of point
Insulating board sheathing, %I”
-ditto plus asphalt coating both sides
Mineral wool (3% inches thick), unprotected
Packaging materials
Cellophane, moisture proof
Glarsine (1 ply waxed or 3 ply plain)
Kraft paper waked with parafin wax, 4.5 Ibs per 100 rq ft
liofilm
Paint Films
2 coots aluminum paint, estimated
2 coats asphalt point, estimated
2 coats lead and oil paint, estimated
2 co& water emulsion, estimated
Papers
Duplex or asphalt lominae
(untreated) 30-30-30, 3.1 lb per 100 sq ft
-ditto 30-60-30,4.2
lb per 100 sq ft
Kraft pope+1 sheet
-2 sheets
-aluminum foil on one side of sheet
--aluminum foil on both sides of sheet
Sheathing paper
Asphalt impregnated and coated, 7 lb per 100 sq ft
Sloterr felt, 6 lb per 100 sq ft, 50% saturated with tar
Roofing Felt, saturated and coated with ospholt
25 lb. per sq ft
50 lb. per sq ft
Tin sheet with 4 holes l/(6 diameter
Crack 12 inches long by ‘/Lo inches wide (approximated from above)
.05
.05
.l
5.0
- .2
- .I
- .6
- 8.0
I
.I5 - .27
.051 - ,091
8.1
5.1
.016
.012
.02 - .lO
1.4
.015
.Ol 1
.17
5.2
/
I
*Pointed surfacer: Two coots of o good vapor seal paint on o smooth surface give o fair vapor barrier. More surface treatment is required on o rough
surface than on o smooth surface. Doto indicates that either asphalt or aluminum paint ore good for vapor seals.
tPluminum Foil on Paper: This material should also be applied over o smooth surface and joints lapped and sealed with asphalt.
vapor barrier should always be placed on the side of the wall having the higher vapor pressure if condensation of moisture in wall is possible.
Application: The heat gain due to water vapor transmission through walls moy be neglected for the normal oir conditioning or refrigeration job.
This latent gain should be considered for air conditioning jobs where there is a great vapor pressure difference between the room and the outside,
particularly when the dewpoint inside must be low. Note that moisture gain due to infiltration usually is of much greater magnitude than moisture
transmission t,hrough building structures.
Conversion Factors: To convert above table values tc: grain/(hr) (sq ft) (inch mercury vapor pressure difference), multiply by 9.8.
grain/(hr) (rq ft) (pounds per sq inch vapor pressure difference), multiply by 20.0.
TO convert Btu latent heat to grains, multiply by 7000/1060= 6.6.
I’,\li’I‘ I. l,O,\D ESTII\l.\TINC
l-86
CONDENSATION OF WATER VAPOR
Whenever there is a difference of temperature and
pressure of water vapor across a structure, conditions
may tlevelop tllat lead to a condensation of moisture.
This condensation occurs at the point of saturation
temperature and pressure.
As water vapor flows thru the structure, its temperature decreases and, if at my point it reaches
the clewpoint or saturation temperature, condensation begins. r\s condensation occurs, the vapor pressure decreases, thereby lowering the dewpoint or
saturation temperature until it corresponds to the
actual temperature. The rate at which condensation
occurs is determined by the rate at which heat is removed from the point of condensation. As the vapor
continues to condense, latent heat of condensation
is released, causing the dry-bulb temperature of the
material to rise.
To illustrate this, assume a frame wall with wood
sheathing.and shingles on the outside, plasterboard
on the inside and fibrous insulation between the
two. Also, assume that the inside conditions are 75 F
db and 50% rh and the outdoor conditions are 0°F
db and 80% rh. Refer to Fig. 28.
The temperature and vapor pressure gradient
decreases approximately as shown by the solid and
dashed lines until condensation begins (saturation
point). At this point, the latent heat of condensation
decreases the rate of temperature drop thru the insulation. This is approximately indicated by the
dotted line.
Another cause of concealed condensation may be
evaporation of water from the ground or damp locations. This water vapor may condense on the under-,
side of the floor joints (usually near the edges where
FIG .
it is coldest) or may llow up thru the outdoor side
of the walls because of stack effect and/or vapor
pressure dilfercnces.
Concealed condensation may cause wood, iron
and brickwork to deteriorate and insulation to lose
its insulating value. These effects may be corrected
by the following methods:
1. Provide unpor bar-?-ien on the high vapor pressure side.
2.
In winter, ventilate the building to reduce the
vapor pressure within. No great volume of air
change is necessary, and normal infiltration
alone is frequently all that is required.
3. In winter, ventilate the structure cavities to
remove vapor that has entered. Outdoor air
thru vents shielded from entrance of rain and
insects may be used.
.
Condensation may also form on the surface of a
building structure. Visible condensation occurs when
the surface of any material is colder than the dewpoint temperature of the surrounding air. In winter,
the condensation may collect on cold closet walls
and attic roofs and is commonly observed as frost on
window panes. Fig. 29 illustrates the condensation
on a window with inside winter design conditions of
70 F db and 40% rh. Point A represents the room
conditions; point B, the dewpoint temperature of
the thin film of water vapor adjacent to the window
surface; and point C, the point at which frost or ice
appears on the window.
Once the temperature drops below the dewpoint,
the vapor pressure at the window surface is also reduced, thereby establishing a gradient of vapor pressure from the room air to the window surface. This
gradient operates, in conjunction with the convec-
28 - CONDENSATION WITHIN FRAME WALL
I;rc.
29
- C ONDENSATION
tive action within the room, to move water Vapor
continuously to the window surface to be condensed,
as long as the concentration of the water vapor is
m
ained in a space.
Visible condensation is objectionable as it causes
staining of surfaces, dripping on machinery and
furnishings, and damage to materials in process of
manufacture. Condensation of this type may be corrected by the following methods:
1. Increase the thermal resistance of walls, roofs
and Hoors by adding kulation with vapor
barriers to prevent condensation within the
structures.
2. Increase the thermal resistance of glass by installing two or three panes with air space(s)
between. In extreme cases, controlled heat,
electric or other, may be applied between the
glass of double glazed windows.
3. Maintain a room dewpoint lower than the lowest expected surface temperature in the room.
4. Decrease surface resistance by increasing the
velocity of air passing over the surface. Decreas,ing the surface resistance increases the window
surface temperature and brings it closer to the
room dry-bulb temperature.
Basis of Chart 2
-Maximum Room
Condensation
RH;
No
Wall,
Roof
or
Glass
Chnrt 2 has been calculated from the equation
u s e d to determine the maximum room dewpoint
temperature that can exist with condensation.
tap = trn, - U(LH - b,,)
fi
where t,, = dewpoint temp of room air, F db
t,.,n = room temp, F
U = transmission coefficient,
Btu/(hr)(sq ft)(deg F)
t,, = outdoor temp, F
ON
WINL~OW S URFACE
fi = inside air film or surEace
Btu/(hr)(sq ft)(deg F)
conductance,
Chad -3 is based upon a room dry-bulb temperature of ‘i0 F db and an inside film conductance of
I.46 Btu/(hr)(sq ft)(deg F).
Use of Chart 2
-Maximum Room
Condensation
RH;
No
Wall,
Roof
or
Gloss
Chart 2 gives a rapid means of determining the
maximum room relative humidity which can be
maintained and yet avoid condensation with a 70 F
db room.
Example ? - Moisture Condensation
Given:
12 in. stone wall with ?h in. sana aggregate
plaster
Room temp - 70 F db
Outdoor temp - 0” F db
Find:
hiaxirnum
room rh without wall condensation.
Solution:
Transmission coefficient U = 0.52 Btu/(hr)(sq
ft)(degF)
(Table 21, page 66)
Maximum room rh = 40.05y0, (Chart 2)
Corrections in room relative humidity for room temperatures other than 70 F clb are listed in the table under
Ciuzrt 2. Values other than those listed may be interpolated.
Example
J0
- Moisture Condensation
Given:
Same as Exorn$e
9, except room temp is 75 F db
Find:
Maximum room rh without wall condensation
Solution:
Transmission coefficient U = 0.52 Btrl/(hr)(sq
ft)(deg F)
(E.uanaple 9)
Xfaximnm room rh for 70 F tlh room temp = 40.05rr/,
(Example 9)
Rh correction for room temp of 75 F tll) with U factor of
0.52 = --1.57rr/, (I,ottom CIlarl 3).
Maximum room rh = 40.05 - 1.57 = 38.48% or 38.5%
1-88
PAI<*
CHART
2-MAXIMUM
ROOM
RELATIVE
HUMIDITY
WITHOUT
I .
LOr\D ES’I‘IM.\~L‘INC;
CONDENSATION
NO WALL, ROOF OR GLASS CONDENSATION
.
- 0
0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
0.9
1.0
WALL, ROOF OR GLASS TRANSMISSION COEFFICIENT “U”
BTU/(HRl(SO
FT)(DEG F)
CORRECTION IN ROOM RH
I.1
(70,
For Wall, Roof or Glass Transmission Coefficient U
u = .65
Outdoor
Temp
IF db)
u = .35
Room Temp (F db)
60
-30
-20
- 1 0
0
10
20
30
40
+2.0
+3.5
+5.0
f7.0
+9.0
+12.0
- 2.0
- 2.5
- 3.5
- 4.0
-7.5
- 9.5
+3.5
f4.0
f5.0
+6.5
f8.5
f9.5
- 3.0
-4.0
- 4.5
- 5.0
- 6.0
-7.5
+2.5yo
f3.0
f3.0
i3.5
f4.0
f4.5
f5.0
f6.0
80
- 2.0 y .
-2.0
- 2.0
- 2.5
- 3.0
- 3.5
- 4.0
- 4.5
1-89
CHAPTER 6. INFILTRATION AND VENTILATION
The data in this chapter is based on ASHAE tests
evaluating the infiltration and ventilation quantities
OE outdoor air. These outdoor air quantities normally have a different heat content than the air
within the conditioned space and, therefore, impose
a load on the air conditioning equipment.
In the case of infiltration, the load manifests itself
directly within the conditioned space. The ventilation air, taken thru the conditioning apparatus,
imposes a load both on the space thru apparatus
bypass effect, and directly on the conditioning
5; :pment.
,. _ he data in this chapter is based on ASHAE tests
and years of practical experience.
INFILTRATION
Infiltration of air and particularly moisture into
conditioned space is frequently a source of sizable
teat gain or loss. The quantity of infiltration air
aries according to tightness of doors and windows,
lorosity of the building shell, height of the buildrg, stairwells, elevators, direction and velocity of
rind, and the amount of ventilation and exhaust
ir. Many of these cannot be accurately evaluated
ad must be based on the judgment of the estimator.
Generally, infiltration may be caused by wind
:locity, or stack effort, or both:
. Wind Velocity - The wind velocity builds up a
pressure on the windward side of the building
and a slight vacuum on the leeward side. The
. outdoor pressure build-up causes air to infiltrate thru crevices in the construction and
cracks around the windows and doors. This, in
turn, causes a slight build-up of pressure inside
the building, resulting in an equal amount of
exfiltration on the leeward side.
!. Difference in Density or Stack Effect - The
variations in temperatures and humidities produce differences in density of air between inside
and outside of the building. In tall buildings
this density difference causes summer and winter infiltration and exfiltration as follows:
Summer - Infiltration at the top and exfiltration at the bottom.
Winter-Infiltration at the bottom and
exfiltration at the top.
This opposite’ direction how balances at some
neutral point near the mid-height of the building. Air flow thru the building openings increases proportionately between the neutral
point and the top and the neutral point and
bottom of the building. The infiltration from
stack effect is greatly influenced by the height
of the building and the presence of open stairways and elevators.
The combined infiltration from wind velocity and
stack effect is proportional to the square root of the
sum of the heads acting on it,
The increased air infiltration flow caused by stack
effect is evaluated< by converting the stack effect
force to an equivalent wind velocity, and tp calculating the flow from the wind velocity data in the
tables.
I
In buildings over 100 ft tall, the equivalent wind
velocity may be calculated from, the following formula, assuming a temperature difference of 70 F db
(winter) and a neutral point at the mid-height of
the building:
Ye = v V” - 1.75a
v, = v vz + 1.75b
where
(for upper
bldgs (for lower
bldgs -
section of
winter)
section of
winter)
_
tall
(1)
tall
(2)
V, = equivalent wind velocity, mph
V = wind velocity normally calculated
for location, mph
a = distance window is above midheight, ft
b = distance window is below midheight, ft
NOTE: The total crackage is considered when
calculating infiltration from stack effect.
INFILTRATION THRU WINDOWS AND DOORS, SUMMER
Infiltration during the summer is caused primarily
by the wind velocity creating a pressure on the windward side. Stack effect is not normally a significant
factor because the density difference is slight, (0.073
lb/cu ft at 7sF db, 50% rh and 0.070 lb/cu ft at
95 F db, 75 F wb). This small stack effect in tall
buildings (over 100 Et) causes air to flow in the top
and out the bottom. Therefore, the air infiltrating
in the top of the building, because of the wind
i
l-90
l’,\l<T
L
I . LOAI)
EST1 ICI;\-I‘ING
Use of Table 41
- Infiltration thru Windows and Doors, Summer
pressure,
tends to flow down thru the building and
out the doors on the street level, thereby olfsetting
some of the infiltration thru them.
In low buildings, air infiltrates thru open doors
on the windward side unless sufficient outdoor air is
introduced thru the air conditioning equipment to
offset it; refer to “0)llsetting Infiltrntion zuith Out-
The data in T(lOle Jf is used to determine the
infiltration thru windows and doors on the windward side with the wind blowing directly at them.
When the wind direction is oblique to the windows
or doors, multiply the values in Tables ?Ila, b, c, d,
by 0.60 and apply to total areas. For specific locations, ntl.just the values in Table 41 to the design
wind velocity: refer to Table 1, Page 10.
door Air.”
With doors on opposite bills, the infiltration can
be considerable if the two are open at the same time.
Basis of Table 41
- infiltration thru Windows and Doors, Summer
During the summer, infiltration is calculated for
the windward side(s) only, because stack effect is
small and, therefore, causes the infiltration air to
flow in a downward direction in tall buiIdings
(over 100 Et). Some of the air infiltrating thru the
windows will exfiltrate thru the windows on the
leeward side(s), while the remaining infiltration air
flows out the doors, thus offsetting some of the infiltration thru the doors. To determine the net infiltration thru the doors, determine the infiltration
thru the windows on the windward side, multiply
this by .80, and subtract from the door infiltration.
For low builclings the door infiltration on the windward side shoulcl be included in the estimate.
The data in Tables -lla, b and c is based on a
wind velocity of 7.5 mph blowing directly at the
window or door, a n d o n o&r_v_eLcrack widths
around typical windows and doors. This data is
‘erivecl from Table 4-/ which lists infiltration thru
cracks around windows and doors as established by
ASHA4E tests.
Table 4Id shows values to be used for doors on
opposite walls for various percentages of time that
each door is open.
The data in Table ille is based on actual tests of
typical applications.
.
TABLE 41 -INFILTRATION THRU WINDOWS AND DOORS-SUMMER*
7.5 mph
T A B L E 4la-DOUBLE H U N G WINDOWS1
Wind Velocity? ’
C F M P E R SQ F T S A S H A R E A
No W-Strip
Average
iABLE
Wood
Poorly
Fitted
Metal
Sash
Sash
Wood
4lb-CASEMENT
L a r g e - 5 4 " x9 6 "
Small--30” x 7 2 "
DESCRIPTION
Sash
W-Strip
Storm
Sash
No W-Strip
W-Strip
.27
.I4
.25
.43
.26
.22
1.20
37
.60
.76
.I7
.24
.80
.35
.40
Sl
.22
Storm
Sash
36
T Y P E WINDOWS$
CFM PER
DESCRIPTION
0%
25%
33%
.72
-
Residential
.33
-
Heavy Projected
-
-
.27
.50
Rolled
Section-Steel
Industrial Pivoted
40%
Hollow Metal-Vertically Pivoted
.39
-
.28
-
-
-
-
.23
.99
.82
Openable
50%
45%
Sash
Architectural Proiected
SO F T S A S H A R E A
Percent
/
-
ROLLED SECTlON STEEL s*w WINDOWS
R~ptESENTr,T,“E
TYPES OF WINDOWS
WIEWED F R O M O”TS,OEI
Area
-
-
.55
.49
.74
-
66Yo
60%
HOLLOY
METAL
WlNDOW
100%
-
2.6
.32
.39
-
1.2
-
2.2
1.45
-
75%
.
.63
.
(:I-l,\I”I‘I-R
6. INl~II.‘I‘I~.\
I‘ION
.\NI)
l-91
Vl~N’I‘II..\‘I‘ION
TABLE 41 -INFILTRATION THRU WINDOWS AND DOORS-SUMMER* (Contd)
7.5 mph Wind Velocityt
TABLE
41c-DOORS O N O N E O R A D J A C E N T W A L L S , F O R C O R N E R E N T R A N C E S
C F M P E R SQ F T A R E A * *
DESCRIPTION
No Use
R e v o l v i n g D o o r s - N o r m a l Operation
Crock
Use
No
Open
Vestibule
Vestibule
.8
5.2
-
-
-
1,200
4.5
10.0
700
500
1 .o
6.5
700
500
Panels Open
Glass Door-%”
Average
CFM
Standing
Wood Door (3’ x 7’)
900
Small Factory Door
Garage 8 Shipping Room Door
Ramp Garage Door
TABLE
4ld-SWINGING D O O R S
ON
OPPOSITE
CFM PER PAIR OF DOORS
y. Time
2nd Door
is Open
% Time 1st Door is Open
.
10
25
50
75
10
100
250
500
750
1,000
25
250
625
1250
1875
2,500
50
75
500
750
1250
1875
2500
3750
5,000
3750
5625
1000
2500
5000
7500
100
I
WALLS
I
100
7,500
/
10,000
I
I
!
I
T A B L E 41e-DOORS
I
APPLICATION
CFM PER PERSON IN ROOM PER DOOR
72“
Revolving
No
I
Bank
Barber
Shop
Candy and Soda
Cigar
Store
Drug
(Small)
Store
Hospital
Swinging
Door
i
Vestibule
6.5
8.0
4.0
5.0
6.0
3.8
5.5
7.0
5.3
20.0
30.6
22.5
6.5
2.0
8.0
6.0
2.5
1.9
5.5
7.0
5.3
1
Department Store
Dress Shop
1..
36”
Door
Room
-
3.5
2.6
Lunch Room
4.0
5.0
3.8
Men’s
Restaurant
2.7
2.0
3.7
2.5
2.8
1.9
Shoe
2.7
3.5
2.6
Shop
Store
*All values in Table 41 are bared on the wind blowing directly at the window or door. When the wind direction is oblique to the window or door, multiply
the above values by 0.60 and use the total window and door oreo on the windward ride(s).
tEased on o wind velocity of 7.5 mph. For design wind velocities different from the base, multiply the above values by the ratio of velocities.
$lncludes
frame
leakage
where
applicable.
**Vestibules may decrease the infiltration os much os 300/,
reducing infiltration.
Example 7 - lnfi/tration
when the door usage is light. When door uroge is heavy, the vestibule is of little
in Tall Buildings, Summer
Given:
A 20-story building in New York City oriented true north.
Building is 100 ft long and 100 ft wide with a floor-to-floor
height of 12 C t . rVal1 area is 5O(r, residential casement
windows having SO(z> fixed sash. There are ten 7 ft x 3 ft
swinging glass doors on the street level facing south.
value for
Find:
Infiltration into the building thru doors and windows,
disregarding outside air thru the equipment and the exhaust
air quantity.
Solution:
The prevailing wind in New York City during the summer
is south, 13 mph (Table 1, page IO).
Correction
l‘he infiltration thru revolving doors is caused by
displacement of the air in the door quadrants, is
almost intlcpendent of wind velocity and, therefore,
cannot be offset by outdoor air.
to Table 1 values for wind velocity
= 1317.5 = 1 . 7 3
Glass arm on south side
= 20 x I2 x 100 x .50 = 12,000 sq ft
Infiltration
thru
windows
Basis of Table 42
- Offsetting Swinging Door Infiltration with Outdoor Air,
Summer
= 12,000 x .49 x 1.73 = 10,200 cfm (Table Jib)
Infltration
thru doors
= IO x 7 X 3 X 10 X 1.73 = 3640 cfm (Table 41~)
Since this building is over 100 ft tall, net infiltration thru
doors = 3640 - (10,200 X .80) = - 4520 cfm.
Therefore,
there is no infiltration thru the doors on the
street level ou tlesijin days, only exfiltration.
OFFSETTING
SUMMER
INFILTRATION
WITH
OUTDOOR
AIR,
Completely offsettin g infiltration by the introduction of outdoor air thru the air conditioning apparatus is normally uneconomical except in buildings
vith few windows and doors. The outdoor air SO
introduced must develop a pressure equal to the
wind velocity to offset infiltration. This pressure
causes exfiltration thru the leeward walls at a rate
equal to wind velocity. Therefore, in a four sided
building with equal crack areas on each side and
the wind blowing against one side, the amount of
outdoor air introduced thru the apparatus must be
a little more than three times the amount that infiltrates. Where the wind is blowing against two sides,
the outdoor air must be a little more than equal to
that which infiltrates.
Offsetting swinging door infiltration is not quite
as difficult because air takes the path of least resistance, normally an open door. Most of the outdoor
air introduced thru the apparatus flows out the door
when it is opened. Also, in tall buildings the window
infiltration tends to flow out the door.
Some of the outdoor air introduced thru the apparatus exfiltrates thru the cracks around the windows and in the construction on the leeward side.
The outdoor air values have been increased by this
amount for typical application as a result of experience.
Use of Table 42
- Offsetting Swinging Door Infiltration with Outdoor Air,
Summer
Table 42 is used to determine the amount of out- .
door air thru air conditioning apparatus required
to offset infiltration thru swinging doors.
Example 2 - Offsetting Swinging Door Infiltration
Given:
A restaurant with 3000 cfm outdoor air being introduced
thru the air conditioning apparatus. Exhaust fans in the
kitchen remove 2000 cfm. Two 7 ft x 3 ft glass swinging
doors face the prevailing wind direction. At peak load conditions, there are 300 people in the restaurant.
.
Find:
The net infiltration thru the outside doors.
Solution: ‘.
Infiltration thru doors = 300 X 2.5 = 750 cfm (Table 41e)
Net outdoor air = 3000 - 2000 = 1000 cfm
Only 975 cfm of outdoor air is required to offset 750 cfm of
door infiltration (Table 42).
Therefore, there will be no net infiltration thru the outside
doors unless there are windows on the leeward side. If
there are windows in the building, calculate as outlined
in Example 1.
TABLE 42-OFFSETTING SWINGING DOOR INFILTRATION WITH OUTDOOR AIR-SUMMER
Net Outdoor Air* (cfm)
Door Infiltration (cfm)
Net Outdoor Air* (cfm)
Door Infiltration (cfm)
140
100
1370
1100
270
200
300
400
500
1480
1560
1670
1760
1200
600
700
800
900
1000
1890
2070
2250
2450
2656
1600
1800
2000
2200
2400
410
530
660
790
920
1030
1150
1260
*Net outdoor air is equal to the outdoor air qwantity introduced thru the apparatus minus the exhaust air quantity.
1300
1400
1500
l-93
CHAPTER r,. INFlLTR,\~I’ION AND VENTIL.\TION
INFILTRATIQN
THRU
WINDOWS
AND
DOORS,
WINTER
Infiltration thru windows and doors during the
winter is caused by the wind velocity and also stack
effect. The temperature clifferenccs
during the winter arc consideral~ly
greater than in summer and,
therefore, the density difference is greater; at 75 F
db nncl SO% rh, tlensity is ,073s; at 0°F db, 40% t-h,
density is .0865. Stack effect causes air to flow in at
the bottom and out at the top, and in many cases
requires spot heating at the doors on the street level
to maintain conditions. In applications where there
is considerable infiltration on the street level, much
of the infiltration thru the windows in the upper
levels will be offset.
Basis of Table 43
- Infiltration thru Windows and Doors, Winter
The data in Table 43 is based on a wind velocity
15 mph blowing directly at the window or door
and on observed crack widths around typical windows and doors. The infiltration thru these cracks
is calculated from Table 44 which is based on
ASHAE tests.
Use of Table 43
- Infiltration thru Windows and Dbors, Winter
Table 43 is used to determine the infiltration of
air thru windows and doors on the windward side
during the winter. The stack effect in tall buildings
increases the infiltration thru the doors and windows
on the lower levels and decreases it on the upper
levels. Therefore, whenever the door infiltration is
increased, the infiltration thru the upper levels must
be decreased by 80% of the net increase in door
infiltration. The infiltration from stack effect on the
leeward sides of the building is determined by using
the difference between the equivalent velocity (V,)
-d the actual velocity (V) as outlined in Example 3.
.le data in Table 43 is based on the wind blowing
directly at the windows and doors. When the wind
direction is oblique to the windows and doors, multiply the values by 0.60 and use the total window and
door area on the windward sides.
Example 3 - Infiltration in Tall Buildings, Winter
Given:
The I>uilcling descrilted in Example 1.
Find:
The infiltration
thru the doors and windows.
Solution:
The prevailing wind in New York City during the winter
is NW at 16.8 mph (Table I, page 10)
Correction on Table 43 for wind velocity is lf.H/15
Since the wind is coming from the Northwest, the
on the north and west sides will allow infiltration
wind is only 603, cffcctive.
Correction
for wind
is 6.
= 1.12.
crackage
Ijut the
direction
Since this huiltling is over 100 Et tall, stack effect causes
inliltration on all sides at the lower Icvels and exfiltration
at the upper levels. The total infiltration on the windward
sitlcs remains the same hecause the increase at the hottom
is exactly equal to the decrease at the top. (For’a floor-hyfloor analysis, use cquivalcnt
wind velocity formulas.) Inliltration thru windows on the windward sides of the lower
levels
= 12,000 X 2 X 1.12 X .G X .98 = 15,810 cfm.
The total inliltration thru the windows on the leeward
sides of the huiltling is equal to the difference hetwcen the
equivalent velocity at the first floor and the design velocity
at the midpoint of the building.
ve =
v-” + 1.751,
II
= 22.2 mph
Ve - V = 22.2 - 16.8 = 5.4 mph
Total
(upper
=
=
infiltration thru windows in lower half
half is exfiltration) on leeward side
12,000 x 2 x IA x (5.4/15) x lh x .98
2160 cfm (Table 43)
of
building
/
NOTE: This is the total infiltration thru the windows on
the leeward side. A floor-by-floor analysis should be
made to balance the system to maintain proper
conditions on each floor.
The infiltration thru the doors on the street level
(on leeward side)
= 10 x 7 x 3 x (5.4/15) x 30
=2310 cfm (Table 43c, average use, 1 and 2
story building).
Example 4 - Offsefting Infiltration with Outdoor Air
Any outdoor air mechanically introduced into the building
offsets some of the infiltration. In Example 3 all of the outdoor air is effective in reducing the window infiltration.
Infiltration is reduced on two windward sides, and the air
introduced thru the apparatus exfiltrates thru the other two
sides.
\
Given:
The building described in Example I with .25 cfm/sq ft
supplied thru the apparatus and 40,000 cfm being exhausted
from the building.
Find:
The net infiltration into this building.
Solution:
Net outdoor air = (.25 X 10,000 X 20) - 40,000 = 10,000 cfm
Net infiltration thru windows
= 15,800 + 2160 - 10,000 = 7970 cfm
Net infiltration thru doors = 2310 cfm (Example 3)
Net infiltration into building = 7970 + 2310 = 10,280 cfm
l-94
PART I. LOAD ESTIM,\TING
TABLE
43-INFILTRATldN THRU WINDOWS AND DOORS-WINTER*
15 mph Wind Velocityt
T A B L E 43a-DOUBLE H U N G W I N D O W S O N W I N D W A R D
SIDE$
CFM PER SO FT AREA
Small-30” x 72”
DESCRIPTION
Average
Poorly
Wood
Fitted
No
Sash
W-Strip
Wood Sash
TABLE
denotes
W-Strip
.52
.42
35
Metol Sash
NOTE:
W-Strip
L a r g e - 5 4 “ x 96”
Storm Sash
2.4
.74
1.60
.69
No
1.2
30
W-Strip
W-Strip
Storm Sash
.53
.33
.26
1.52
.47
.74
1.01
.44
.50
weatherstrip.
43b-CASEMENT T Y P E
WINDO W
S
O N
WINDW
A R D SIDEI
C F M P E R SQ FT AREA
.
Percent
DESCRIPTION
25%
0%
Roiled
Section-Steel
.65
-
Architectural Projected
Residential
Hollow
4oYo
450/o
50%
Area
60%
66Yo
7%
~w%
5.2
Sarh
Industrial Pivoted
Heavy
33Yo
Ventilated
Projected
Metal-Vertically
Pivoted
1.44
-
1.98
-
.78
-
-
-
-
-
.45
.98
-
-
-
1.64
-
-
-
-
-
.56
-
.54
1.19
-
1.1
-
2.9
-
I.48
-
-
-
-
-
1.26
.63
.7a
-
-
4.3
2.4
.
T A B L E 43c-DOORS O N O N E O R A D J A C E N T W I N D W A R D
SIDES1
CFM PER
SQ FT AREA**
Average Use
DESCRIPTION
Infrequent
Use
I
Revolving
Door
Glass
D o o r - f % / Crack)
Wood Door 3’
Small
Garage B Shipping Room Door
Romp Garage Door
10.5
30.0
2.0
I
/
I
1.6
9.0
x 7’
Factory Door
1.5
4.0
4.0
Tall Building (ft)
182
Story Bldg.
13.0
!
13.0
50
1
12.6
36.0
I
14.2
40.5
15.5
I
100
17.3
49.5
17.5
I
200
21.5
I
9.0
13.5
*All values in Table 43 are based on the wind blowing directly at the window or door. When the prevailing wind direction is oblique to the window
or doors, multiply the above values by 0.60 and use the total window and door oreo on the windward side(s).
tBased on a wind velocity of 15 mph. For design wind velpcities different from the base, multiply the table values by the ratio of velocities.
$Stock effect in toll buildings may also cause infiltration on the leeward ride. To evaluate this, determine the equivalent velocity (Ve) and subtract the
design velocity (V). The equivalent velocity ir:
VC = \I Vz- 1.750 (upper section)
Ve=$ Vi-l- 1.75b (lower section)
Where a and b are the distances above and below the mid-height of the building, respectively, in ft.
Multiply the table valuer by the ratio (V.,- V)/15 for doors and one half of the windows on the leeward side of the building. (Use valuer under
“1 and 2 Story Bldgt” for doors on leeward side of toll buildings.)
**Doors
on opposite sides increase the above values 25y* Vestibules may decrease the infiltration 03 much (IS 3Oyo when door uroge is light. If door
usage is heavy, the vestibule is of little value in reducing infiltration. Heat added to the vestibule will help maintain room temperature near the door.
(:H,\l’~I‘ICK
fi. INI~Il.~I‘K,\‘I‘ION
INFILTRATION
.\NI)
l-95
VEN’I‘II.,\‘I‘ION
Use of Table 44
- CRACK METHOD (Summer or Winter)
- Infiltration thru Windows and Doors, Crack Method
The crack method ol evaluating infiltration is
more accurate than the arca methods. It is difficult
Tnble # is used to determine the infiltration thru
the doors and windows listed. This table does not
take into account winter stack effect which must be
evaluated separately, using the equivalent wind
velocity formulas fireviously presented.
to establish the exact crack dimensions but, in certain close tolerance applications, it may be necessary
to evaluate the load accurately. The crack method is
applicable both summer and winter.
Infiltration thru Windows, Crack Method
Example 5 Basis of Table 44
- Infiltration thru Windows and Doors, Crack Method
Given:
A 4 ft x 7 ft rcsitlcntial
The data on windows in Table 44 are based on
ASHAE tests. These test results have been reduced
20% because, as infiltration occurs on one side, a
certain amount of pressure builds up in the building, thereby reducing the infiltration. The data on
glass and factory doors has been calculated from
observed typical crack widths.
Fintl:
The inliltration
casement wintlow facing south.
thru this window.
Solution:
Assume the crack widths are measured as follows:
Wintlow frame - none, well sealed
Wintlow openable area - l/32 in. crack; length. 20 ft
Assume the wintl velocity is 30 mph due south.
Infiltration thru winclow = 20 X 2.1 = 42 cfm (Table 44)
4
TABLE
44--INFILTRATION
THRU
WINDOWS
AND
DOORS-CRACK METHOD-SUMMER-WINTER*
I
TABLE 44o-DOUBLE
HUNG WINDOWS-UNLOCKED ON WINDWARD SIDE
CFM PER LINEAR FOOT OF CRACK
3
Wind
TYPE OF
DOUBLE HUNG WINDOW
5
No W- WStrip
Strip
Wood Sash
Average Window
Poorly Fitted Window
Poorly Fitted--with Storm Sash
Mekd S a s h
Tba1.E
44b-CASEMENT
.12
.45
.23
.33
Velocity--Mph
15
10
No W Strip
WStrip
.07
.35
.lO
1.15
.05
.lO
.57
.7a
.22
.32
.16
.32
20
No W- W Strip
Strip
.65
1 .a5
.93
1.23
25
No W- W Strip
Strip
.40
.57
.29
.53
.98
2.60
1.30
1.73
30
No W- W Strip
Strip
.I50
.85
.43
77
1.33
3.30
1.60
2.3
.82
1.18
.59
1.00
No W- W Strip
Strip
1.73
4.20
2.10
2.8
1.05
1.53
.76
1.27
TYPE WINDOWS ON WINDWARD SIDE
CFM PER LINEAR FOOT OF CRACK
TYPE OF CASEMENT WINDOW AND
TYPICAL CRACK SIZE
Wind
Velocity-Mph
5
10
15
20
25
30
‘b’s” crock
‘h” crock
‘A” crack
.a7
.25
.33
1.80
.60
.87
2.9
1.03
4.1
1.43
1.47
1.93
5.1
1.06
2.5
6.2
2.3
3.0
Residential Casement
R e s i d e n t i a l Casement
‘!&“ crack
‘h” crack
.lO
.23
.30
.53
.55
.07
1.27
1 .oo
1.67
1.23
2.10
Heavy Casement Section Projected
Heavy Casement Section Projected
‘%A” crack
‘%I” crock
.05
.13
.17
.40
.30
.63
.43
.90
.58
1.20
30
1.53
1.46
2.40
3.10
3.70
4.00
R o l l e d Section-Steel Sash
Industrial Pivoted
Architectural Projected
Architectural Projected
Hollow
Metal-Vertically Pivoted
.
.50
‘Infiltration caused by stack effect must be calculated separately during the winter.
*No allowance has been made for usage. See Table 43 for infiltration due to usage.
\
.7a
TABLE 44-INFILTRATIONTHRU WINDOWS AND DOORS-CRACK METHOD-SUMMER-WINTER*
(Contd)
TABLE 44c-DOORSt
ON WINDWARD SIDE
CFM PER LINEAR FOOT OF CRACK
TYPE OF DOOR
Wind Velocity- mph
5
10
15
20
25
30
3.2
4.8
6.4
6.4
1go
13.0
9.6
14.0
19.0
13.0
20.0
26.0
16.0
24.0
26.0
19.0
29.0
38.0
.45
.90
.90
.60
1.2
2.3
.90
1.8
3.7
1.3
2.4
5.2
1.7
3.3
6.6
2.1
4.2
8.4
6.4
9.6
13.0
16.0
19.0
G l o s s Door-Herculite
Good Installation ‘h “ crack
Average Installation ?&‘I crock
Poor
Installation
l/i” crock
Ordinary Wood or Metal
Well Fitted-W-Strip
Well Fitted-No W-Strip
Poorly Fitted-No W-Strip
Factory Door ‘/‘a”
crock
3.2
/
.
VENTILATION
VENTILATION
STANDARDS
The introduction of outdoor air for ventilation of
conditioned sp?ces is necessary to dilute the odors
given off by people, smoking and other internal
air contaminants.
The amount of ventilation required varies
primarily with the total number of people, the ceiling height and the number of people smoking.
People give off body odors which require a minimum of 5 cfm pel person for satisfactory dilution.
J an and one half cfm per person is recommended.
This is based on a population density of 5J to 75
J s-r person and a typical ceiling height of 8 ft.
With greater population densities, the ventilation
quantity should be increased. When people smoke,
the additional odors given off by cigarettes or cigars
.equire a minimum of 15 to 25 cfm per person. In
special gathering rooms with heavy smoking, 30 to
50 cfm per person is recommended.
Basis of Table 45
-Ventilation
Standards
The data in Table Ji is based on test observation
of the clean outdoor air required to maintain satisfactory odor levels with people smoking and not
smoking. These test results were then extrapolated
for typical concentrations of people, both smoking
and not smoking, for the applications listed.
Use of Table 45
-Ventilation
Standards
Table 1’5 is used to determine the minimum and
recommended ventilation air quantity for the listed
a
applications. In applications where the minimum
values are used and the minimum cfm per person
and cfm per sq It of Noor area are listed, use the
larger minimum quantity. Where the crowd density
is greater than normal or where better than satisfactory conditions are desired, use the recommended
values.
*
SCHEDULED
VENTILATION
In comfort applications, where local codes permit,
it is possible to reduce the capacity requirements of
the installed equipment by reducing the ventilation
air quantity at the time of peak load. This quantity
can be reduced at times of peak to, in effect, minimize the outdoor air load. At times other than peak
load, the calcuiated outdoor air quantity is used.
Scheduled ventilation is recommended only for installations operating more than 12 hours or 3 hours
longer than occupancy, to allow some time for flush-
ing out the building when no odors are being generated. It has been found, by tests, that few complaints of stuffiness are encountered when the outdoor air quantity is reduced for short periods of
time, provided the flushing period is available. It is
recommended that the outdoor air quantity be reduced to no less than 40y0 of the recommended
quantity as listed in TnOle 45.
The procedure for estimating and controlling
scheduled ventilation is as follows:
1. In estimating the cooling load, reduce the air
quantity at design conditions to a minimum of
40y0 of the recommended air quantity.
2. Use a dry-bulb thermostat following the cooling and dehumidifying apparatus to control
the leaving dewpoint such that:
\
(:H,\I’TEK
fi. INFILTKA’I‘ION i\Nl)
l-97
VENTIL;\‘rION
a. With the dewpoint at design, the damper
motor closes the outdoor air damper to 40Cj’0
of the design ventilation air quantity. .
b. As the dewpoint decreases below design, the
outdoor air damper opens to the tlesign
setting.
Example 6 -
Solution:
The population density is typical, 100 sq ft per person, but
the number of snlokers is considerable.
Recommended ventilation = 50 X 15 = 750 cfm (Tnble 45)
Minimum
Ventilation Air Quantity, Office Space
Given:
A 5000 sq ft office with a ceiling height of 8 ft and 50 people.
Approximately 4070 of the people smoke.
Find:
The
ventilation
air
ventilation
= 50 X 10 = 500 cfm (Tnble $5)
500 cfm will more than likely not main&in
satisfactory
conditions within the space because the number of smokers
is considerable. Therefore, 750 cfm should be used in this
apphcation.
NOTE: Many applications have exhaust fans. This means
that the outdoor air quantity must at least equal
the exhausted air, otherwise the infiltration rate
will increase. Tables -/6 and 47 list the approximate
capacities of typical exhaust fans. The data in these
tables were obtained from published ratings of several manufacturers of exhaust fans.
quantity.
TABLE 45-VENTILATION STANDARDS
CFM
APPLICATION
Banking Space
S M O K I N G
.
15
25
15
10
Occosionol
Very Heavy
Corridors (Supply or Exhaust)
t
Department Stores
Directors Rooms
Drug Stores t
Heavy
-
50
30
-
N o n e
Extreme
50
7%
Five and Ten Cent Storer
NO”=
Funeral Parlors
NO”.?
-
GO’O&
Some
Meeting Rooms
VeryHeavy
Restaurant ‘Ofeteriat
Dining Room t
Some
NO”=
Considerable
Considerable
Considerable-
School Rooms $
N o n e
Shop Retail
N0’le
Theate’S
NO”8
Theater
Some
.33
-
7%
30
25
5
30
.25
.05
-
7%
10
30
20
Heavy
I
/‘/
/
10
N o n e
None
None
I _ --
7%
10
I
CFM PER
SQ F T O F F L O O R
Minimum*
IO
Considerable
None
Factories$$
R e s i d e n c e
20
30
10
Occasional
Broker’s Board Rooms
Cocktail Bars
Laboratoriest
Minimum*
Considerable
Beauty Parlors
PERSON
Recommended
Some
Some
I
Barber Shops
Operating RoomsS**
Hospitals Private Rooms
t Words
Hotel Rooms
I
PER
I
I
1
-
30
-
25
15
25
-
20
15
50
30
IO
15,
25
10
12
5
25
30
7%
1.25
.29
.25
-
Toil&f (Exhaust)
*When minimum is used, use the larger.
fSee local codes which moy govern.
tMay be governed by exhaust.
@Jse these values unless governed by other sources of contamination or by local codes.
**All outdoor air is recommended to overcome explosion hazard of anesthetics.
\
I’/\KT I. LOAD ESTlhI.\-I‘ING
1-98
TABLE 47-PROPELLER FAN CAPACITIESFREE DELIVERY
TABLE 46-CENTRIFUGALFAN
CAPACITIES
Motor
Ou1let
Horrepower
Range
Velocity
Ranaa (fD”l)
Fan Diameter
(ill.)
a
Speed
(rpm)
Capacity*
kfm)
1500
500
l/70-1 /20
800-2000
l/20-1/6
500-2500
12
12
l/20-%
MO-2900
l/5-2
950-4300
1000-2000
18
850
1800
1000-2000
18
1140
2350
1000-2000
20
850
2400
1000-2000
20
1140
2750
20
1620
3300
*These typical air capacities were obtained from published ratings of
several manufacturers of notionally known exhaust fans, single width,
single inlet. Range of static pressures 1% to 1 ‘A inches. Fans with inlet
diameter 10 inches and smaller are direct connected.
*The capacity of these fans has been arbitrarily taken at 1000 fpm
minimum and 2000 fpm maximum outlet velocity. Far there fans the
usual selection probably is approximately 1500 fpm outlet velocity
for ventilation.
1140
025
1725
1100
16
055
1000
16
1140
1500
*The capacities of fans of various manufacturers may
from the values given above.
vary
+
10%
.
II
l-99
CHAPTER 7. INTERNAL AND SYSTEM HEAT GAIN
INTERNAL HEAT GAIN
Internal heat gain is the sensible and latent heat
released within the air conditioned space by the
occupants, lights, appliances, machines, pipes, etc.
This chapter outlines the procedures for determining the instantaneous heat gain from these sources.
A portion of the heat gain from internal sources
is radiant heat which is partially absorbed in the
building structure, thereby reducing the instantaneous heat gain. Chapter 3, “Heat Storage, Diversity and Stratification,” contains the data and
methods for estimating the actual cooling load from
.’ 7 heat sources referred to in the following text.
PEOPLE
Heat is generated within the human body by
oxidation, commonly called metabolic rate. The
metabolic rate varies with the individual and with
his activity level. The normal body processes are
performed most efficiently at a deep tissue temperature of about 98.6 F; this temperature may vary only
thru a narrow range. However, the human body is
capable of maintaining this temperature, thru a
wide ambient temperature range, by conserving or
dissipating the heat generated within itself.
This heat is carried to the surface of the body by
the blood stream and is dissipated by:
1. Radiation from the body surface to the surrounding surfaces.
2. Convection from the body surface and the respiratory tract to the surrounding air.
. Evaporation of moisture from the body surface
and in the respiratory tract to the surrounding
air.
The amount of heat dissipated by radiation and
convection is determined by the difference in temperature between the body surface and its surroundings. The body surface temperature is regulated by
the quantity of blood being pumped to the surface;
the more blood, the higher the surface temperature
up to a limit of about 96 F. The heat dissipated by
evaporation is determined by the difference in vapor
pressure berween
the body and the air.
Basis of Table 48
- Heat Gain from People
Table 48 is based on the metabolic rate of an average adult male, weighing 150 pounds, at different
levels of activity, and generally for occupancies
longer than S hours. These have been adjusted for
typical compositions of mixed groups of males and
females for the listed applications. The metabolic
rate of women is about 85% of that for a male, and
for children about 75yo.
The heat gain for restaurant applications has
been increased 30 Btu/hr sensible and 30 Btu/hr
latent heat per person to include the food served.
The data in Table 48 as noted are for continuous
occupancy. The excess heat and moisture brought
in by people, where short time occupancy is occurring (under 15 minutes), may increase the heat gain
from people by as much as 10%.
Use of Table 48
- Heat Gain from People
To establish the proper heat gain, the room design
temperature and the activity level of the occupants,
must be known.
Example 1 - Bowling Alley
Given:
A 10 lane bowling alley, 50 people, with a room design
dry-bulb temperature of 75 F. Estimate one person per
alley bowling, 20 of the remainder seated, and 20 standing.
Find:
Sensible and latent heat gain from people.
Solution:
Sensible heat gain = (10 X 525) + (20 X 240) + (20 X 280)
= 15,650 Rtu/hr
Latent heat gain = (10 X 925) + (20 x 160) + (20 X 270)
= 17,850 Btu/hr
LIGHTS
Lights generate sensible heat by the conversion of
the electrical power input into light and heat. The
heat is dissipated by radiation to the surrounding
surfaces, by conduction into the adjacent materials
and by convection to the surrounding air. The
radiant portion of the light load is partially stored,
and the convection portion may be stratified as
described on page 39. Refer to Table 12, page 35, to
determine the actual cooling load.
Incandescent lights convert approximately 10%
of the power input into light with the rest being
generated as heat within the bulb and dissipated by
radiation, convection and conduction. About 80yo
of the power input is dissipated by radiation and
only about 10% by convection and conduction,
Fig. 30.
.
3 1 -CONVERSIONOFELECTRICPOWERTO
F IG .
H EAT
AND
LIGHT WITH FLUORESCENT LIGHTS,
,~PI'ROXI~IATE
F1c.30 -CONVERSION OF ELECTRICPOWERTO
HEAT ANDLIGHTWITHJNCANDESCENTLIGHTS,
/~I'I'ROxIMATE
Fluorescent lights convert about 25y0 of the power
input into light, with about 25y0 being dissipated
by radiation to the surrounding surfaces. The other
50% is dissipated by conduction and convection. In’
addition to this, approximately 257’, more heat is
generated as heat in the ballast of the fluorescent
lamp, Fig. 31.
Table 49 indicates the basis for arriving at the
gross heat gain from fluorescent or incandescent
lights.
.
TABLE 4B-HEAT GAIN FROM PEOPLE
Aver-
DEGREE OF
ACTIVITY
Seated at rest
Seated, very
work
Office
Theater,
Grade
High
walking
Walking,
School
390
seated
350
175
175
195
155
210
140
230
120
260
90
School
450
400
Offices,
Apts.,
Hotels,
College
180
220
195
205
215
185
240
160
275
125
475
-
Dept., Retail, or
Variety Stoic
450
180
270
200
250
215
235
2.F
205
285
165
5 5 0
Drug Store
=j
Standing,
slowly
1
light
worker
Standing,
slowly
TYPICAL
APPLICATION
“ROOM
DRY-BULB
TEMPERATURE
age
M e t - Adobolic
justed
80 F ‘,
78 F
75 F ’
02 F
70 F
Rate
Met( A d u l t abolic
Btu/hr
Btu/hr
Btu/hr
Btu/hr
Btu/hr
Rate*
Mole)
I
Btu/hr Btu/hr Sensible Latent Sensible Latent Sensible Latent Sensible ( Latent Sensible Latent
500
Bank
550
Sedentary work
Rertaurantt
500
590
light
Factory.
800
900
bench
Moderate
Walking,
1
180
320
1
200
300
/
220
280
1
255,
245
1
290
210
walking
work
dancing
3
mph
Dance
light
work
Hall
Factory, fairly
heavv work
1000
190
360
220
330
240
310
280
270
320
230
750
190
560
220
530
245
505
295
455
365
385
850
220
630
245
605
275
575
325
525
400
450
1000
270
730
300
700
330
670
380
620
460
540
605
845
I
Heavy work
B o w l i n g Alleyt,
Factory
1500
1450
4 5 0 ’ 1000
*Adjusted Metabolic Rote is the metabolic rote to be applied to o
mixed group of people with o typical percent composition based on
the following foctorr:
Metabolic rate, adult femole=Metabolic
rote, adult male X 0.85
Metabolic rote, children
=Metabolic
rote, adult mole X 0.75
I
465
985
/
485
965 /
I
525
921
tllestouront-Values
for thir application include 60 Btu per hr for food
per Individual (30 8tu reralble and 30 Btu latent heat per hrl.
fBowling-Assume
one person per alley actually bowling and oil others
sitting, metabolic rote 400 Btu per hr; or standing, 550 Btu per hr.
I
i !
(;H/\I”I’EI<
,‘<
TYPE
HEAT GAIN* Btu/hr
Fluoresc+
Total Light Watts X i.‘%T X 3.4
Incandescent
Total Light Watts X 3.4
*Refer to Tables 12 and 13, pager 35-37 to determine actual cooling
load.
tfluorescent
light wattage is multiplied by 1.25 to include heat gain
in ballast.
APPLIANCES
iLfost appliances contribute both sensible and
latent heat to a space. Electric appliances contribute
latent heat, only by virtue of the function they
l-101
,-,’
7 . INTERN,\I./IN11 SYS’I‘ICXI Flli.\~I‘ (;.\IN
TABLE 49-HEAT GAIN FROM LIGHTS
,’
perform, that is, drying, cooking, etc., whereas gas
burning appliances contribute additio4al moisture
as a product of combustion. A properly designed
hood with a positive exhaust system removes a considerable amount of the generated heat and moisture
from most types of appliances.
Basis of Tables 50 thru 52
- Heat Gain from Restaurant
Miscellaneous
Appliances
Appliances
and
The data in these tables have been determined
from manufacturers data, the American Gas Association data, Directory of Approved Gas Appliances
and actual tests by Carrier Corporation.
TABLE 50-HEAT GAIN FROM RESTAURANT APPLIANCES
NOT
OVERALL
DIMENSIONS
Less Legs and
Handles (In.)
APPLIANCE
HOODED*-ELECTRIC
TYPE
OF
CONTROL
Coffee Brewer--‘/i gal
Man.
Wormer-!/i g a l
Man.
4 Coffee Brewing Units
MISCELLANEOUS
DATA
RECOM HEAT GAIN’
MAINMFR
FOR AVG USE
TAIN- r
MAX
ING
Sensible Latent
Total
RATING
RATE
Heat
Heat
Heat
Btu/hr
Btu/hr
Btu/hr
Btu/hr
Btu/hr
2240
306
25x 30 x 26H
Auto.
Water heater-2000
watts
Brewers-2960 WOWS
16900
Coffee Urn-3 gal
- 3 gal
-5 gal
1 5 D i o x 34H
12 x 23 oval x 2 1 H
18 Dio x 37H
Man.
Auto.
Auto.
Black finish
Nickel plated
Nickel plated
11900
15300
17000
Doughnut
22 x 22 x 57H
Auto.
Exhaust system t o
W ho motor
16000
with 4% g a l T a n k
Machine
outdoors-
10 x 13 x 25H
Egg Boiler
Food
Warmer
with
Plate
Warmer, per sq ft top
surface
Auto.
\od Warmer without
Plate W a r m e r , per sq ft
top
Man.
Auto.
surface
Med. ht.-550 watts
._ L o w h t - 2 7 5 watts
Insulated, separate
heating unit for each
pot. Plate warmer in
base
Ditto, without plate
wormer
Fry Kettle-l 1% lb fat
12 Dia x 14H
Auto.
Fry Kettle’-25 lb fat
16xlBx12H
Auto.
Frying area 12” x 14”
Griddle,
18 x 18 x 8H
Auto.
F r y i n g t o p 1 B” x 14”
G r i l l e , Meqt -
14 x 14 X~ iOH
Auto.
Cooking orea 10” x 12”
Grille,
13 x 14 x 10H
Auto.
Roll Warmer (
26xl7x13H
Toaster,
15 x 15 x 2BH
Toaster,
-
Frying
.
Sandwich
Continuous
Continuous
Toaster, P o p - U p
Waffle
Iron
,
Waffle Iron for Ice Cream
Sandwich
.
1 3740
306
306
3000
2600
3600
900
230
220
90
1120
320
4800
1200
6000
2600
2200
3400
1700
1500
2300
4300
3700
5700
5000
1
1 1200
5aon
/
BOO
1 2000
1350
500
350
350
700
1020
400
200
350
550
8840
1100
1600
2400
4000
23800
2000
3800
5700
9500
8000
3100
1700
4800
10200
2800
*'
1900
3900
2100
6000
Grill area 12” x 12”
5600
1900
2700
700
3400
Auto.
One
1500
400
1100
100
1200
Auto.
2 Slices wide360 slices/hr
7500
5000
5100
1300
6400
drawer
20 x 15 x 2BH
Auto.
4 Slices wide720 slicer/hr
10200
6000
6100
2600
8700
6xllx9H
Auto.
2 Slices
4150
1000
2450
450
2900
12 x 13 x 10H
Auto.
One waffle 7” dia
2480
600
1100
750
1850
14 x 13 x 10H
Auto.
12 Cakes,
7500
1500
3100
2100
* I f p r o erly d e s i g n e d p o s i t i v e e x h a u s t h o o d i s u s e d , m u l t i p l y r e c o m m e n d e d
P
each 2%” x 3sqn
value by .50.
I
5200
,.
l-102
PAR-I‘
I. I.O/\D
ES-I‘ihl,\‘l‘ING
.
Use of Tables 50 thru 52
- Heat Gain from Restaurant
Miscellaneous
Appliances
Appliances
and
The M~~intnining
Rate is the heat generated when
the appliance is being maintained at operating temperature but not being used.
The Recommended for Avel-nge Use values arc
those which the appliance generates under normal
use. These appliances seldom operate at maximum
capacity during peak load since they are normally
warmed up prior to the peak.
The values in Tables 50 thru 52 are for unhooded
appliances. If the appliance has a properly designed
positive exhnust hood, reduce the sensible and the
latent heat gains by 50%. A hood, to be effective,
should extend beyond the appliance approximately
4 inches per foot of height between the appliance
and the face of the hood. The lower edge should not
bc higher than 4 feet above the appliance and the
average fact velocity across the hood should not be
less than 70 fpm.
TABLE 51 -HEAT GAIN FROM RESTAURANT APPLIANCES
NOT HOODED*-GAS BURNING AND STEAM HEATED
OVERALL
DIMENSIONS
Less Legs and
Handles (In.)
APPLIANCE
TYPE
OF
CONTROL
GAS
Coffee Brewer-% gal
wormer--/i gal
Mon.
Man.
RECOM HEAT GAIN
FOR AVG USE
MAINMFR
TAINMAX
ING
RATING
RATE
Btu/hr
Btu/hr
MISCELLANEOUS
DATA
.
Sensible Lotent
Heat
Heat
Btu/hr
Btu/hr
Total
Heat
Btu/hr
BURNING
Combination brewer
and wormer
3400
500
500
3200
3900
4 Brewers and 4%
gal tank
1350
400
350
100
1700
500
7200
1BOO
9000
2900
2900
5800
Coffee Brewer Unit with
Tank
19 x 30 x 26H
C o f f e e u r n - 3 gal
15” Dia x 34H
Auto.
Block finish
Coffee U r n - 3 g a l
1 2 x 23 oval x 2lH
Auto.
Nickel plated
3400
2500
2500.
5000
Coffee u r n - 5 g a l
18 D i a x 37H
Auto.
Nickel plated
4700
3900
3900
7800
/
Food Warmer, Values
SQ . fl too surface
per
I Man. I
W a t e r b o t h t;pe
I-
,
2000
/
900
Fry Kettle-15 lb fat
12 x 20 x 1BH
Auto.
Frying area 10 x 10
14250
3000
Fry Kettle-28 lb fat
15 x 35 x 1lH
Auto.
Frying orea 11 x 16
24000
4500
22 x 14 x.17H
s (1.4 $4 ft
-grill surface)
Man.
Insulated
2 2 , 0 0 0 Btu/hr
1 5 , 0 0 0 Btu/hr
37000
.3 1
Man.
Ring type burners
12000 to 22000
Btu/ea
Grill-Brqil-O-Grill
Top Burner
Bottom Burner
:
Stoves, Short OrderOpen Top. Valuer per
tq ft top surface
Stoves, Short OrderC l o s e d T o p . Valuer per
sq ft top surface
Toaster,
Continuous
.
15 x 15 x 2BH
Man.
Auto.
STEAM
(
850
1
450
14400-1 3600
/
1300
1 BOO0
1
14000
4200
4200
8400
Ring type burners 10000
to 12000 Btu/ea
11000
3300
3300
6600
2 Slices wide360 slices/hr
12000
7700
3300
11000
10000
HEATED
C o f f e e u r n - 3 gal
- 3 gal
- 5 gal
1 5 D i o x 34H
12 x 23 oval x 21H
18 Dia x 37H
Auto.
Auto.
Auto.
Black finish
Nickel plated
Nickel plated
2900
2400
3400
1900
1600
2300
4800
4000
5700
Coffee
1 5 D i a x 34H
12 x 2 3 oval x 21H
18 Dia x 37H
Man.
Man.
Man.
Black finish
Nickel plated
Nickel plated
3100
2600
3700
3100
2600
3700
6200
5200
7400
500
900
Urn-3 g a l
- 3 gal
-5 g a l
Food Warmer, per sq ft
t o p surface
Auto.
400
Food Warmer, per sq ft
top surface
Man.
450
*If properly designed positive exhaust hood is used, multiply recommended value by .50.
1150
1500
(;~-l/\l”1‘1111
7 .
IN’I‘EI~N,\I,
,\Nl> SYS~I‘I-~c1
l-103
I rb:.\‘i (;,\I&
TABLE 52-HEAT GAIN FROM MISCELLANEOUS APPLIANCES
NOT HOODED*
TYPE
OF
CONTROL
APPLIANCE
MFR
MAX
RATING
MISCELLANEOUS DATA
RECOM HEAT GAIN FOR AVG USE
Sensible
Heat
Btu/hr
Btu/hr
totent
Heat
Btu/hr
Total
Heat
Btu/hr
ELECTRIC
Hair Dryer, Blower Type
15 amps, 115 volts AC
Man.
Fan 165 watts,
(low 915 wat+s, high 1580 watts)
5,370
2,300
400
2,700
Hair Dryer, helmet type,
6.5 amps, 115 volts AC
Man.
Fan 80 watts,
(low 300 watts, high 710 watts)
2,400
1,870
330
2,200
Mon.
60 heaters at 25 watts each,
36 in normal ure
5,100
150
1,000
23,460
35,460
Permanent Wave Machine
i
Pressurized
Instrument
Washer and Sterilizer
1l”X Il”X 2 2 ”
Neon Sign, per
linear ft tube
‘I$” outside dia
%” outside dio
Solution and/or
Blanket
Warmer
18” x 30” x 72”
18” x 24” x 72”
Steriliker
Dressing
Sterilizer,
Auto.
Auto.
Rectangular
Sterilizer,
Bulk
Auto.
Auto.
Auto.
Auto.
Auto.
Auto.
Auto.
Water
’
850
12,000
30
60
1,200
1,050
16” x 24”
20” x 36”
24” x
24” x
24” x
~~~~~-x-
24”
24”
36”
3.6”
10 gallon
15 gallon
3,000
2,400
4,200
3,450
8,700
24,000
18,300
47,300
34,800
4 1,700
56,200
68,500.
161,700 /
184,000
2 10,000
2 1,000
27,000
36,000
45,000
97,500
140,000
180,000
55,800
68,700
92,200
113,500
259,200
324,000
390,000
4,100
6,100
16,500
24,600
9,600
23,300
x 36”
x 48”
x 48”
x-m
36” x 42” x 84”.
42” x 48” x 96”
48” x 54” x 96”
Auto.
Auto.
30
60
Sterilizer,
Instrument
Auto.
Auto.
Auto.
Auto.
Auto.
6” x 8” x 17”
9” x 10” x 20”
10” x 12” x 22”
10”~ 12”~ 3 6 ”
12” x 16” x 24”
2,700
5,100
8,100
10,200
9,200
Sterilizer,
Utensil
Auto.
Auto.
16” x 16” x 24”
20” x 20” x 24”
10,600
12,300
20,400
25,600
Auto.
Auto.
Model 120 Amer Sterilizer Co
Model 100 Amer Sterilizer Co
2,000
1,200
4,200
2,100
Sterilizer,
Hot
Air
Water Still
5 gal/hour
X-ray Machines, for
making pictures
Physicians and Dentists office
’ X-ray Machines,
for therapy
Burners,
Laboratory
small bunsen
small bunsen
fishtail
burner
fishtail burner
large bunsen
Cigar
Lighter
Dryer System
5 helmets
10 helmets
I
20,600
30,700
2,400
3,900
5,900
9,400
8,600
1,700
5,100
9,000
14,000
19,600
17,800
3 1,000
37,900
6,200
3,300
2,700
NOW
4,400
NOW
NOW
Heat load may be opprecioblewrite mfg for data
GAS
BURNING
Mon.
36 dia barrel with
manufactured gas
1,800
Man.
Mon.
56 dia with nat gos
%6 dia with nat gas
Mon.
Mon.
36 dio bar with nat gas
I ‘/2 dio mouth, adi orifice
Mon.
Continuous flame type
Auto.
Auto.
Consists of heater L fan which blows
hot air thru duct system to
helmets
Hair
‘If properly designed positive exhaust hood is used, multiply recommended value by .50
1
960
240
3,000
3,500
1,680
1,960
420
490
2,100
2,450
5,500
6,000
3,080
3,350
770
850
3,850
4,200
2,500
1
900
/
100
4,000
6,000
j
1
1,200
1,000
19,000
27,000
’
l-104
PART
Example 2 - Restaurant
Given:
A restaurant with the following electric appliances with a
properly designed positive exhaust hood on each:
1. Two 5.gallon coffee urns, both used in the morning,
only one used either in the afternoon or evening.
2. One 20 sq Et food warmer without plate warmer.
3. Two 24 x 20 x 10 inch frying griddles.
4. One 4.slice pop-up toaster, used only in the morning.
5. Two 25 II) deep fat, fry kettles.
Find:
Heat gain from
evening meal.
Solution:
Use Table
these
appliances
during
the
50.
afternoon
Sensible
1. Coffee Urn - only one in use:
Sensible heat gain = 3400 X .50 =
Latent heat gain = 2300 X .50 =
1700
2. Food Warmer:
Sensible heat gain = 20 X 200 X .50 =
Latent heat gain = 20 x 350 x 50 =
2000
3.
and
Latent
1150
3500
Frying Griddles:
Sensible heat gain = 2 X 5300 X .50 =
Latent heat gain = 2 x 2900 x .50 =
1.
L O A D ESTIM,\TINC
lf the fluid is conveyed outside of the air conditioned space, only the inefficiency of the motor
driving fan or pump should be included in room
sensible heat gain.
If the temperature of the fluid is maintained by
a separate source, these heat gains to the fluid heat
of compression are a load on this separate source
only.
The heat gain or loss from the system sho~~lci be
calculated separately (“System Heat Gain,” p. 120).
Motors driving process machinery (lathe, punch
press, etc.): The total power input to the machine
is dissipated as heat at the machine. If the product
is removed from the conditioned space at a higher
temperature than it came in, some of the heat
input into the machine is removed and should not
be considered a heat gain to the conditioned space.
.
The heat added to a product is determined by multiplying the number of pounds of material handled
per hour by the specific heat and temperature rise.
5300
2900
4. Toaster - not in use
5. Fry Kettles:
Sensible heat gain = 2 X 3800 X .50 =
Latent heat gain = 2 x 5700 x .50 =
3800
5700
Total sensible heat gain =
12,800
Total latent heat gain =
13,250
L
ELECTRIC MOTORS
Electric motors contribute sensible heat to a
space by converting the electrical power input to
heat. Some of this power input is dissipated as heat
in the motor frame and can be evaluated as
input X (1 - motor eff).
The rest of the power input (brake horsepower
?r motor output) is dissipated by the driven machine
and in the drive mechanism. The driven machine
utilizes this motor output to do work which may or
may not result in a heat gain to the space.
Motors driving funs and pumps: The power input
increases the pressure and velocity of the fluid and
the temperature of the fluid.
The increased energy level in the fluid is degenerated in pressure drop throughout the system and
appears as a heat gain to the fluid at the point where
pressure drop occurs. This heat gain does not appear
as a temperature rise because, as the pressure reduces, the fluid expands. The fluid expansion is a
cooling process which exactly offsets the heat
generated by friction. The heat of compression required to increase the energy level is generated at
the fan or pump and is a heat gain at this point.
.
Basis of Table 53
- Heat Gain from Electric Motors
Table 53 is based on average efficiencies of squirrel
cage induction open type integral horsepower and
fractional horsepower motors. Power supply for
fractional horsepower motors is 110 or 220 Golts, 60
cycle, single phase; for integral horsepower motors,
208, 220,‘or 440 volts, 60 cycle, 2 or 3 phase general
purpose and constant speed, 1160 or 1750 rpm. This
table may also be applied with reasonable accuracy
to 50 cycle, single phase a-c, 50 and 60 cycle enclosed
and fractional horsepower polyphase motors.
Use of Table 53
- Heat Gain from Electric Motors
The data in Table 53 includes the heat gain from
electric motors and their driven machines when
both the motor and the driven machine are in the
conditioned space, or when only the driven machine
is in the conditioned space, or when only the motor
is in the conditioned space.
Caution: The power input to electric motors does
not necessarily equal the rated horsepower divided by the motor efficiency.
Frequently these motors may be operating
under a continuous overload, or may be
operating at less than rated capacity. It is
always advisable to measure the power
input wherever possible. This is especially important in estimates for industrial
installations where the motor-machine
load is normally a major portion of the
cooling load.
CHAI’TER
l-105
7 . INTERNAL ANIl SYSTI:M HE,\‘I’ G,\Ih-
When retdings me obtairred dizrtly in watts and
when both motors and driven machines 3x-c in the
air conditioned space, the heat gain is equal to the
number of watts times the factor 3.4 Btu/(watt)(hr).
When the machine is in the contlitioncd space
and the motor outside, multiply the watts by the
motor efficiency ant1 by the factor 3.4 to determine
heat gain to the space.
When the machine is outside the conditioned
multiply the watts by one minus the motor
efficiency antI by the Factor 3.4.
SIX’CC,
~\lthough the results are less accurate, it may be
expedient to obtain power input measurements
using ;I clamp-on ammeter and voltmeter. These
instruments permit instantaneous readings only.
They afford means for determining the load factor
but the usage factor must be obtained by a careful
investigation of the operating conditions.
TABLE 53-HEAT GAIN FROM ELECTRIC MOTORS
CONTINUOUS
OPERATION*
LOCATION OF EQUIPMENT WITH RESPECT TO
C O N D I T I O N E D S P A C E O R A I R SlREAMt
NAMEPLATE+
OR
B R A K E
HORSEPOWER
Btu per Hour
‘I20
‘I12
‘vi
‘ii
‘yi
40
49
55
60
64 \
320
430
580
710
1,000
130
210
320
430
640
190
220
260
280
360
66
70
72
79
80
1,290
1,820
2.680
3,220
4.770
850
1,280
1,930
2,540
3,820
440
540
750
680
950
2
3
5
?‘/i
10
80
01
82
85
85
6,380
9,450
15,600
22,500
30,000
5,100
7,650
12,800
19,100
25,500
1,280
1,800
2,800
3,400
4,500
15
20
25
30:
40
06
87
88
09
89
44,500
58,500
72,400
85,800
1 15,000
38,200
51,000
63,600
76,400
102,000
6,300
7,500
8,800
9,400
13,000
50
60
75
100
125
89
09
90
90
90
143,000
172,000
212,000
284,000
354,000
127,000
153,000
191,000
255,000
3 18,000
16,000
19,000
2 1,000
29,000
36,000
150
200
250
91
91
91
420,000 ’
560,000
700,000
382,000
5 10,000
636,000
38,000
50,000
64,000
m
1%
‘h
%
1
1 ‘/a
.
I
‘For intermittent operation, an appropriate usage factor should be used, preferably measured.
tlf motors are overloaded and amount of overloading is unknown, multiply the above heat gain factors by the following maximum service fadorr:
Maximum Service Factors
Horsepower
AC Open Type
DC Open Type
‘/20-H
1.4
-
‘h-‘/i
1% - ?4
1
1 N-2
3-250
1.35
-
I .25
-
1.25
1.15
1.20
1.15
1.15
1.15
No overload is allowable with enclosed motors.
SFor a fan or pump in air conditioned space, exhausting air and pumping fluid to outside of space, use values in last column.
l-106
I’AKT
I .
LOAD ESTliV,\TING
\
following is a conversion table which can be
used to determine load Eactors from measurements:
The
TO FIND
+
HP
OUTPUT
KILOWATTS
INPUT
Direct
Current
IXEXeff
I X E
746
1,000
1
IXEXpfXeff
IXEXpf
Phase
746
1,000
3 or 4 Wire
I X E X pf X elf X 1.73
I X E X pf X 1.73
3 Phase
746
1,000
4 Wire
IXEXpfXeffXZ
IXEXZXpf
2 Phase
746
Where I = amperes
E = volts
IOTE:
I
eff = efficiency
pf = power factor
For 2 phase, 3 wire circuit, common conductor current
is 1.41 times that in either of the other two conductors.
Example
3 - Electric Motor Heaf Gain in a Factory
(Motor 8hp Established by a Survey)
Given:
1. Forty-five 10 hp motors operated at 80’9” rated capacity,
driving various types of machines located within air
conditioned space (lathes, screw machines, etc.)
Five 10 hp motors operated at 800/, rated capacity, driving
screw machines, each handling 5000 11~s of bronze per
hr. Both the final product and the shavings from the
screw machines are removed from the space on conveyor belts. Rise in bronze temperature is 30 F; sp ht is
.Ol Btu/(lb) (F).
2. Ten 5 hp motors (5 bhp) driving fans, exhausting air to
the outdoors.
3. Three 20 hp motors (20 bhp) driving process water
pumps, water discarded outdoors.
Find:
Total heat gain from motors.
8olution:
Use Table 53.
Sensible Heat
Btu/hr
1. Machines - Heat gain to space
= 45 X 30,000 X .80 =
1,080,OOO
Heat gain from screw machines
= 5 X 30.000 X .80 = 120,000 Btu/hr
Heat removed from space from
screw machine work
= 5000 X 5 X 30 X .Ol = 7,500 Btu/hr
Net heat gain from screw machines
to space
= 120,000 - 7,500 =
112,500
2. Fan exhausting air to the outdoors:
Heat gain to space = 10 X 2800 =
28,000
3. Process water pumped to outside
air conditioned space
Heat gain to space = 3 x 7500 =
22,500
Total heat gain from motors on
machines, fans, and pumps =
1,243,OOO
NOTE: If the process water were to be recirculated and
coolctl in the circuit from an outside source, the
heat gain to the water
3 X (58,500 - 7500) = 153,000 Btu/hr
would I,ecomc a load on this outside source.
!
!
!
/
:‘I
PIPING, TANKS AND EVAPORATION OF WATER FROM
A FREE SURFACE
1,000
I
!
1
Gain
Hot pipes and tanks add sensible heat to a space
by convection and radiation. Conversely, cold pipes
remove sensible heat. All open tanks containing hot
water contribute not only sensible heat but also
latent heat due to evaporation.
In industrial plants, furnaces or dryers are often
encountered. These contribute sensible heat to the
space by convection and radiation from the outside
surfaces, and frequently dryers also contribute sensible and latent heat from the drying process.
Basis of Tables 54 thru 58
- Heat Gain from Piping, Tanks and Evaporation
of Water
Table 54 is based on nominal flow in the pipe and
a convection heat flow from a horizontal pipe of 1.016 x(&)2x
x
(&).lX’
(temp diff between hot water or steam
.
and room).
The radiation from horizontal pipes is expressed
bY 17.23 x lo-10 x emissivity x (T,” - T,“)
where T, = room surface temp, deg R
T, = pipe surface temp, deg R
Tables 55 and 56 are based on the same equation
and an insulation resistance of approximately 2.5
per inch of thickness for 85yo magnesia and 2.9 per
inch of thickness with moulded type.
Caution: Tables 55 and 56 do not include an allowance for fittings. A safety factor of 10%
should be added for pipe runs having
numerous fittings.
Table 57 is based on an emissivity of 0.9 for
painted metal and painted or bare wood and concrete. The emissivity of chrome, bright nickel plate,
stainless steel, or galvanized ‘iron is 0.4. The resistance (r) of wood is 0.833 per inch and of concrete
0.08 per inch. The metal surface temperature has
been assumed equal to the water temperature.
NOTE: The heat gain from furnaces and ovens can
be estimated from Table 57, using the outside temperature of furnace and oven.
Table 58 is based on the following formula for
still air: Heat of evaporation = 95 (vapor pressure
)
/
/
.i
.
l-107
CHAPTER 7 . I N T E R N A L :\NI) S Y S T E M H E A T GAIN
differential between water and air), where vapor
pressure is expressed in inches of mercury, and the
room conditions are 75 F db and 50% rh.
,
b u l b temperature (Ib/hr x t e m p diff x .45). T h e
latent heat gain is equal to the pounds per hour
escaping times 1050 Btu/lb.
Use of Tables 54 thru 58
- Heat Gain from Piping, Tanks and Evaporation
of Water
MOISTURE
ABSORPTION
When moisture (regain) is absorbed by hygroscopic materials, sensible heat is added to the space.
The heat so gained is equal to the latent heat of
vaporization which is approximately 1050 Btu/lb
times the pounds of water absorbed. This sensible
heat is an addition to room sensible heat, and a
deduction from room latent heat if the hygroscopic
material is removed from the conditioned space.
Example 4 - Heat Gain from Hot Water Pipe and
Storage Tank
Given:
Room conditions - 75 F db, 5070 rh
50 ft of IO-inch uninsulated hot water (125 F) pipe.
The hot water is stored in a 10 ft wide x 20 ft long x 10 ft
high, painted metal tank with the top open to the atmosphere. The tank is supported on open steel framework.
LATENT HEAT GAIN-CREDIT
SENSIBLE HEAT
Find:
Sensible and latent heat gain
Btu/hr
Jing - Sensible heat gain = 50 X 50 X 4.76 =
11,900
Tank - Sensible heat gain, sides
= (20 x 10 x 2) + (10 x 10 x 2)
X50X1.8=
54,000
- Sensible heat gain, bottom
= (20 x 10) x 50 x 1.5 =
15,000
ROOM
Some forms of latent heat gain reduce room
sensible heat. Moisture evaporating at the room
wet-bulb temperature (not heated or cooled from
external source) utilizes room sensible heat for heat
of evaporation. This form of latent heat gain should
be deducted from room sensible heat and added to
room latent heat. This does not change the total
room heat gain, but may have considerable effect on ,
the sensible heat factor.
When the evaporation of moisture derives its heat
from another source such as steam or electric heating
coils, only the latent heat gain to the room is figured;
room sensible heat is not reduced. The power input
to the steam or electric coils balances the heat of
evaporation except for the initial warmup of the
water.
Solution:
Use Tables 54,57 and 58
TO
Total sensible heat gain =
80,900
Total latent heat gain, top = (20 X 10) X 330 = 66,000
STEAM
When steam is escaping into the conditioned
space, the room sensible heat gain is only that heat
represented by the difference in heat content of
steam at the steam temperature and at the room dry-
TABLE 54-HEAT TRANSMISSION COEFFICIENTS FOR BARE STEEL PIPES
Btu/(hr) (linear ft)
H O T
-dOMINAL
PIPE
SIZE
(in.)
??H
3
3H
4
5
6
0
10
12
’
(deg F diff between pipe and surrounding air)
W A T E R
300 F
100 psig
338 F
157 F
230 F
268 F
0.61
0.75
0.92
1.14
1.29
0.71
0.87
1.07
1.32
1.49
0.76
0.93
1.51
1.58
2.15
2.43
2.72
2.26
2.55
2.85
1.84
2.19
2.63
2.97
3.32
1.99
2.36
2.84
3.22
3.59
3.30
3.89
4.96
6.09
7.15
3.47
4.07
5.21
6.41
7.50
4.05
4.77
6.10
7.49
a.80
4.39
5.16
6.61
a.12
9.53
120 F
150 F
18OF
50 F
80
110 F
140 F
0.46
0.56
0.68
0.85
0.96
0.50
0.61
0.74
0.92
1.04
0.55
0.67
0.82
1.01
1.15
0.58
0.72
0.88
1.09
1.23
1.28
1.53
1.83
2.06
2.30
1.41
1.68
2.01
2.22
2.53
2.80
3.29
4.22
5.18
6.07
3.08
3.63
4.64
5.68
6.67
-‘.
1.18
1.40
1.68
1.90
2.12
2.58
3.04 '
3.88
4.76
5.59
* A t 7 0 F d b room temperature
_
5 psig
227 F
210 F
TEMPERATURE
F
STEAM
I
5 0 psig
DIFFERENCE*
I.80
I
1.aa
1.15
1.43
1.63
l-108
PART 1. LOAD ESTIMATING
L
TABLE 55-HEAT TRANSMISSION COEFFICIENTS FOR INSULATED PIPES*
Btu/(hr)
(linear ft) (deg F diff between pipe and room)
IRON PIPE
SIZE
(in.)
1 In. Thick
1% In. Thick
2 In. Thick
%
0.16
%
0.18
0.20
0.14
0.15
0.17
0.12
0.13
0.15
1 ‘/!I
1 ‘/I
2
2%
0.24
0.26
0.30
0.35
0.20
0.21
0.24
0.27
0.17
0.18
0.2 I
0.24
3
31%
4
5
0.40
0.45
0.49
0.59
0.32
0.35
0.38
0.45
0.27
0.30
0.32
0.38
6
0.68
0.85
1.04
1.22
0.52
0.65
0.78
0.90
0.43
0.53
0.64
0.73
a5 PERCENT MAGNESIA INSULATIONt
1
a
10
12
.
*No allowance for fittings. This table applies only to straight runs of pipe. When
numerous fittings exist, a suitable safety factor must be included. This added heat
gain at the fittings moy be OS much as 10%. Generally this table can be used
without adding this safety factor.
tOther insulation. if other types of insulation we used, multiply the above
values by the factors shown in the following table:
PIPE
MATERIAL
COVERING
Corrugated Asbestos (Air Cell)
4 Ply per inch
6 Ply per inch
8 Ply per inch
Laminated Asbestos (Sponge Felt)
Mineral Wool
Diatomaceous
Silica
(Super-X)
Brown Asbestos Fiber (Wool Felt)
FACTORS
1.36
1.23
1.19
0.98
1 .oo
1.36
0.88
TABLE 56--HEAT TRANSMISSION COEFFICIENTS FOR INSULATED COLD PIPES*
MOULDED TYPEt
Btu/(hr) (linear ft) (deg F diff between pipe and room)
.,
,.“,,,.”
,.
v’““;I/
‘,I
ICE WATER
litbN:tiIPE,SiZE.
.,:.”
(in.) 1,’
.,.”
1”:1
_)_
:,.;;;
‘/?2
:’ 3/! ‘.
:‘:;y,
‘,
..~ “;..::.. ,i i ‘A, ,: T.
: :’ 1 ‘A
,,,I,”
2.
1
:
,,
‘*
.2’/a
{‘I,
BRINE
.,v+ ;:,HEAyY j B R I N E ‘
i
Coefficient
Actual Thickness
of Insulation (In.1
Coefficient
0.11
0.12
0.14
2.0
2.0
2.0
0.10
0.1 I
0.12
2.8
2.9
3.0
0.09
0.09
0.10
1.6
1.5
1.5
1.5
0.16
0.17
0.20
0.23
2.4
2.5
2.5
2.6
0.13
0.13
0.15
0.17
3.1
3.2
3.3
3.3
0.1 1
0.12
0.13
0.15
1.5
1.5
I .7
1.7
0.27
0.29
0.30
0.35
2.7
2.9
2.9
3.0
0.19
0.19
0.2 I
0.24
3.4
3.5
3.7
3.9
0.16
0.18
1.7
0.40
3.0
0.26
4.0
0.23
II .9
.9
1.9
0.46 0.56
0.65
3.0 3.0
3.0
0.32 0.38
0.44
4.0 4.0
4.0
0.26 0.31
0.36
Actual Thickness
of Insulation (In.)
1.5
1.6
1.6
.il
Actual Thicknesi. :, :* :. ! _I’
of Insulation (In.)
Coefficient
0.18
0.20
*No allowance for fittings. This table applier only to straight runs of pipe. When numerous fittings exist, a suitable safety factor must be included. This
added heat goin at the fittings moy be OS much ~1s 10%. G enerally this table can be used without adding this safety factor.
tlnrulation material. Valuer in this table are based on CI material having o conductivity kE0.30.However, o 15% safety factor war added to this k
value to compensate for seams and imperfect workmanship. The table applies to either cork covering (kE0.291, or mineral wool board (k=0.32). The
thickness given above is for molded mineral wool board which is usually some 5 to 10% greater than molded cork board.
CH/II’TEK 7 .
IN-I’EKN;\I. .\NI>
TABLE
SYS’I‘ICRI
FJ1:.\‘1‘
57-HEAT TRANSMJSSION
l-109
C..\IN
COEFFICIENTS FOR UNINSULATED TANKS
SENSIBLE HEAT GAIN*
Btu/(hr) (sq ft) (deg F diff between liquid and room)
METAL
Painted
CONSTRUCTION
Temp
Vertical(Sider)
TOP
l3attmn
Bright (Nickel)
Diff
50 F
100 F
150 F
1.8
2.1
1.5
2.0
2.4
1.7
2.3
2.7
2.0
Temp
WOOD
2 % in. Thick
CONCRETE
6 in. Thick
Painted or Bare
Painted or Bore
Diff
Temp
Tamp Diff
Diff
*To estimate latent heat load if water is being evaporated, see Table 58
TABLE 58-EVAPORATION FROM A FREE WATER SURFACE-LATENT HEAT GAIN
STILL AIR, ROOM AT 75 F db.
WATER TEMP
Btu/(hr)(sq
ft)
i
75 F
42
1
I
50% RH
100 F
1
125 F
1
150 F
1
175 F
1
200 F
140
/
330
)
680
(
1260
)
2190
SYSTEM HEAT GAIN
The system heat gain is considered as the heat
added to or lost by the system components, such as
the ducts, piping, air conditioning fan, and pump,
etc. This heat gain must be estimated and included
in the load estimate but can be accurately evaluated
only after the system has been designed.
SUPPLY AIR DUCT HEAT GAIN
The supply duct normally has 50 F db to 60 F
dl
r flowing through it. The duct may pass
through an unconditioned space having a temperature of, say, 90 F db and up. This results in a heat
gain to the’ duct before it reaches the space to be
conditioned. This, in effect, reduces the cooling
capacity of the conditioned air. To compensate for
it, the cooling capacity of the air quantity must be
increased. It is recommended that long runs of ducts
in unconditioned spaces be insulated to minimize
heat gain.
Basis or Chart 3
- Percent Room Sensible Heat to be Added for Heat
Gain to Supply Duct
Chart 3 is based on a difference of 30 F db between supply air and unconditioned space, a supply
duct velocity of 1800 fpm in a square duct, still air
on the outside of the duct and a supply air rise of 17
F db. Correction factors for different room temperatures, duct velocities and temperature differences are
included below Chnrt 3. Values are plotted for use
with uninsulated, furred and insulated ducts.
Use of Chart 3
- Percent Room Sensible Heat to be Added for Heat
Gain to Supply Duct
To use this chart, evaluate the length of duct
running thru the unconditioned space, the temperature of unconditioned space, the duct velocity, the
suppIy air temperature, and room sensible heat subtotal.
Example 5 - Heat Gain to Supply Duct
Given:
20 ft of uninsulated duct in unconditioned space at 100 F dh
Duct velocity - 2000 fpm
Supply air temperature - 60 F db
Room sensible heat gain - 100,000 Btu/hr
Find:
Percent addition to room sensible heat
Solution:
The supply air to unconditioned
ence = 100 - GO = 40 F db
From Chart 3, percent
Correction for 40 F
2000 fpm duct velocity
Actual percent addition
space
temperature
addition = 4.5%
db temperature difference and
= 1.26
= 4.5 X 1.26 = 5.7%
differ-
I’.\K-r I. I.O.\I) I:S’I‘lhI.\~I‘ING
l-110
CHART 3-HEAT GAIN TO SUPPLY DUCT
Percent
of
Room
Sensible
Heat
0
0
1000
2000
DUCT
3000
4000
5000
V E L O C I T Y (FPMI
.
MULTIPLYING FACTORS FOR
OTHER ROOM TEMPERATURES
Room Temp
Q
= UPI X
(2.162;‘6x5
Multiplying
75
1.10
76
77
78
79
80
1.06
1 .oo
0.97
0.94
Factor
.
0.92
/r&A: UPI (‘3--11)
where:
A = duct area (sq ft)
Q = duct heat gain (Btu/hr)
U = duct heat transmission factor (Btu/hr-sq
ft-F)
V = duct velocity (fpm)
P = rectangular duct perimeter (ft)
tl = temperature of supply air entering duct
I = duct length (ft)
t3
Based on formulas in ASHRAE
SUPPLY
AIR
DUCT
LEAKAGE
LOSS
Air leakage from the supply duct may be a serious
loss of cooling effect, except when it leaks into the
conditioned space. This loss of cooling effect must
be added to the room sensible and latent heat load.
Experience indicates that the average air leakage
from the entire length of supply ducts, whether large
or small systems, averages around lo<;& of the supply
air quantity. Smaller leakage per foot of length for
larger perimeter ducts appears to be counterbalanced by the longer length of run. Individual work-
(F)
= temperature of surrounding air (F)
Guide 1963, p. 184, 185.
manship is the greatest variable, and duct leakages
from 5yo to 30% have been found. The following is
a guide to the evaluation of duct leakages under
various conditions:
I. Bare ducts within conditioned spice - usually
not necessary to figure leakage.
2. Furred or insulated ducts within conditioned
space - a matter of judgment, depending on
whether the leakage air actually gets into the
room.
(;H,\I’~1‘~:11
7 .
IN’I‘EIIN,\I, :\NI) SYS’I‘EM
1-111
HE,\‘1 C;,\IN
TABLE 59-HEAT GAIN FROM AIR CONDITIONING FAN HORSEPOWER, DRAW-THRU SYSTEMIf
CENTRAL STATION SYSTEMS$
APPLIED OR UNITARY SYSTEM**
Temp Diff
Room to Supply Air
Temp Diff
Room to Supply Air
10 F
15 F
20 F
Fan Motor
Not in
Conditioned
Space
or
Air Stream
F a n Motortt
in
Conditioned
space
or
Air Stream
1.25
3.9
1so
4.6
L----l-
1.75
2.00
30 F
25 F
PERCENT
OF
ROOM
5.00
6.00
8.00
19.2
24.4
38.0
12.8
16.3
25.4
9.6
12.2
19.0
7.7
9.9
15.2
6.4
0.2
12.7
0.75
1 .oo
1.6
2.6
3.6
1.1
1.8
2.4
0.8
1.3
1.8
0.6
1.1
1.5
0.5
0.9
1.2
2.7
4.2
5.8
1.8
2.8
3.8
5.0
6.0
3.4
4.0
7.0 4.7
2.5
3.0
3.5
2.0
2.4
2.8
1.7
2.0
2.4
7.6
9.2
10.7
5.1
6.1
7.2
1.25
1
so
I
2.00
3.00
4.00
30 F
5.4
4.00
0.50
L
25 F
HEAT*
6.2
10.4
15.3
3.00
20 F
15 F
10 F
SENSIBLE
1.75
I
1.4
2.1
2.9
I
1.1
1.7
2.3
I
0.9
1.4
1.9
8.0
13.2
19.0
*Excludes from heat gain, typical values for bearing losses, etc. which are dissipated in apparatus room.
tFon Total Pressure equals fan static pressure plus velocity pressure at fan discharge. Below 1200 fpm the fan total pressuce
the fan static. Above 1200 fpm the total pressure should be figured.
$7Oyo
is approximately equal to
fan efficiency assumed.
**5Oyo fan efficiency assumed.
tt8Oyo
motor and drive efficiency assumed.
$$For draw-thru systems, this heat is on addition to the supply air heat gain and is added to the room’ sensible heat. For blow-thru systems this fan heat
is added to the grand total heat; use the RSH times the percent listed and add to the GTH.
3. All ducts outside the conditioned space assume 10% leakage. This leakage is a total
loss and the full amount must be included.
When only part of the supply duct is outside
the conditioned space, include that fraction of
10% as the leakage. (Fraction is ratio of length
outside of conditioned space to total length of
supply duct.)
HEAT GAIN FROM AIR
FAN HORSEPOWER
CONDITIONING
The inefficiency of the air conditioning equip=
ment fan and the heat of compression adds heat
to the system as described under “Electric Motors.”
In the case of draw-through systems, this heat is an
addition to the supply air heat gain and should be
added to the room sensible heat. With blow-through
systems (fan blowing air through the coil, etc.) the
fan heat added is a load on the dehumidifier and,
therefore, should be added to the grand total heat
(see “Percent Addition to Grand Total Heat”).
Basis of Table 59
- Heat Gain from Air Conditioning Fan Horsepower
The air conditioning fan adds heat to the system
in the following manner:
1. Immediate temperature rise in the air due to
the inefficiency of the fan.
2. Energy gain in the air as a pressure and/or
velocity rise.
3. With the motor and drive in the air stream or
conditioned space, the heat generated by the
inefficiency of the motor and drive is also an
immediate heat gain.
The fan efficiencies are about 707, for central
station type fans and about 50% for packaged
equipment fans.
1-112
PAKT
Use of Table 59
- Heat Gain from Air Conditioning Fan Horsepower
The approximate system pressure loss and dehumidified air rise (room minus supply air temperature) differential must be estimated from the system
characteristics and type of application. These should
be checked from the final system design.
The normal comfort application has a dehumidified air rise of between 15 F db and 25 F db and the
fan total pressure depends on the amount of ductwork involved, the number of fittings (elbows, etc.)
in the ductwork and the type of air distribution
system used. Normally, the fan total pressure can be
approximated as follows:
1. No ductwork (packaged equipment) - 0.5 to
1.00 inches of water.
2. Moderate amount of ductwork, low velocity
systems - 0.75 to 1.50 inches of water.
3. Considerable ductwork, low velocity system 1.25 to 2.00 inches of water.
4. Moderate amount of ductwork, high pressure
system - 2.00 to 4.00 inches of water.
5. Considerable ductwork, high pressure system
- 3.00 to 6.00 inches of water.
Example 6 -Heat Gain from Air Conditioning Fan
Horsepower
Given:
Same data as Example 5
80 ft of supply duct in conditioned space
Find:
Percent addition to room sensible heat.
Solution:
Assume 1.50 inches of water, fan total pressure, and
20 F db dehumidifier rise. Refer to Table 59.
Heat gain from fan horsepower = 2.3’%
SAFETY FACTOR AND PERCENT ADDITIONS TO ROOM
SENSIBLE AND LATENT HEAT
A safety factor to be added to the room sensible
heat sub-total should be considered as strictly a
factor of probable error in the survey or estimate,
and should usually be between 0% and 5yo.
The total room sensible heat is the sub-total plus
percentage additions to allow for (1) supply duct
heat gain, (2) supply duct leakage losses, (3) fan
horsepower and (4) safety factor, as explained in the
preceding paragraph.
Example 7 - Percent Addition to Room Sensible Heat
Given:
Same data as Examples 5 and 6
Find:
Percent addition to room sensible heat gain sub-total
.
I. LOAD ESTIMATING
Solution:
=
Supply duct heat gain
Supply duct leakage (20 ft duct of total 100 ft) =
=
Fan horsepower
=
Safety factor
5.7%
2.07,
2.37”
0.07,
=
10.0%
Total percent addition to RSH
The percent additions to room latent heat for
supply duct leakage loss and safety factor should be
the same as the corresponding percent additions to
room sensible heat.
RETURN AIR DUCT HEAT AND LEAKAGE GAIN
The evaluation of heat and leakage effects on
return air ducts is made in the same manner as for
supply air ducts, except that the process is reversed;
there is inward gain of hot moist air instead of
loss of cooling effect.
.
Chart 3 can be used to approximate heat gain to
the return duct system in terms of percent of RSH,
using the following procedure:
1. Using RSH and the length of return air duct,
use Chart 3 to establish the percent heat gain.
2. Use the multiplying factor from table below
Chart 3 to adjust the percent heat gam for
actual temperature difference between the air
surrounding the return air duct and the air inside the duct, and also for the actual velocity.
3. Multiply the resulting percentage of heat gain
by the ratio of RSH to GTH.
4. Apply the resulting heat gain percentage to
GTH.
To determine the return air duct leakage, apply
the following reasoning:
1. Bare duct within conditioned space - no inleakage.
2. Furred duct within conditioned space or furred
space used for return air - a matter of judgment, depending on whether the furred space
may connect to unconditioned space.
3. Ducts outside conditioned space - assume up
to 3’7, inleakage, depending on the length of
duct. If there is only a short connection between conditioned space and apparatus, inleakage may be disreg?rded. If there is a long
run of duct, then apply judgment as to the
amount of inleakage.
HEAT GAIN FROM DEHUMIDIFIER PUMP HORSEPOWER
With dehumidifier systems, the horsepower required to pump the water adds heat to the system as
outlined under “Electric Motors”. This heat will
be an addition to the grand total heat.
1-113
TABLE 60-HEAT GAIN FROM DEHUMIDIFIER PUMP HORSEPOWER
L A R GE PUMPS+
SMALL PUMPS* O-100 GPM
CHILLED
5 F
7 F
WATER
TEMP
10 F
RISE
12 F
PUMP HEAD
(ftl
35
70
100
*Efficiency
50%
CHILLED
15 F
5 F
7 F
100 GPM AND LARGER
WATER
TEMP
RISE
10 F
12 F
15 P
0.5
1.5
2.0
0.5
I .o
I .5
0.5
I .o
I .o
PERCENT OF GRAND TOTAL HEAT
2.0
3.5
5.0
1.5
2.5
4.0
tEfficiency
1 .o
2.0
2.5
I .o
1.5
2.0
0.5
1 .o
1.5
1.5
2.5
4.0
!
1 .o
2.0
3.0
70%
Basis of Table 60
- Heat Gain from Dehumidifier Pump Horsepower
Table 60 is based on pump efficiencies of 50% for
small pumps a n d 70% Eor large pumps. Small
pumps are considered to have a capacity of less than
100 gallons; large pumps, more than 100 gallons.
u.
I Table 60
- Heat Gain from Dehumidifier Pump Horsepower
The chilled water temperature rise in the dehumidifier and the pump head must be approximated
to use Table 60.
1. Large systems with considerable piping and
fittings may require up to 100 ft pump head;
normally, 70 ft head is the average.
2. The normal water temperature rise in the dehumidifier is between 7 F and 12 F. Applications using large amounts of water have a lower
rise; those using small amounts of water have
a higher rise.
PERCENT
ADDITION
TO
GRAND
TOTAL
HEAT
The percent additions to the grand total heat
to compensate for various external losses consist of
heat and leakage gain to return air ducts, heat gain
from the dehumidifier pump horsepower, and the
heat gain to the dehumidifier and piping system.
These heat gains can be estirpated as follows:
1. Heat and leakage gain to return air ducts, see
above.
2. Heat gain from dehumidifier pump horsepower, Table 60.
3. Dehumidifier and piping losses:
a. Very little external piping - 1% of GTH.
b. Average external piping - 2% of GTH.
c. Extensive external piping - 4% of GTH.
4. Blow-through fan system - add percent room
sensible heat from Table 59 to GTH.
5. Dehumidifier in conditioned apparatus room reduce the above percentages by one half.
I
1-115
CHAPTER 8. APPLIED PSYCHROMETRICS
The lxcccclingchapters contztin the
t o lxol)erly evnlu:tte the he:tting ant1
They also reconlnlcnd ottt(loor air
ventilzttion l>url~oses i n :tre;ts where
locnl cocks tlo not exist.
2 . Air u~rttli~ionir~~ :~)+wmtus - ktctors Afccting
cot~t~non
lxocesses xntl the clfcct oC these lactors
on selection of xir contlitionitig ccluilxncrtt.
prncticnl clata
cooling loacls.
quantities for
state, city or
II/ pa7~tinl lorrtl co77 trd - t h e
clfect ol lxtrtial lo;~ct on ecluilxncnt s e l e c t i o n
and on the cotntnon lxocesses.
3. P.ryrl~7~omeI~ic.r
T h i s challter tlcscrilxx lxxticallxychrometrics
as al~l~lietl to ~tl~lxttxttts selection. It is tlivitlctl into
three parts:
1.
Descliplior7 of m ’ w r s , p7wesses
encoitnterec!
cntions.
and factors
7’0 hells rccognizc ternis, lactors z~ntl lxocesses
dcscrilxxl in this chapter, a brief tlefinition of lxychronletrics is offeretl at this point, along with an
illusttxtion aticl tlcfinition 0E terms nl>l>earing on ;1
stanclartl lxychrotnetric chart (Fig. 32).
- as
in nortnztl air contlitioning appli-
rv.bulb Temperature -The temperature of air as registered by
.n orditiazy thermometer.
Specific
dry air.
Temperature -The temperature registered by a thermometer whose bulb is covered by a wetted wick and exposed
to a current of rapidly moving air.
Wet-bulb
Temperature-The temperature at which
tion of moisture begins when the air is cooled.
Dewpoint
Volume
-The cubic feet of the mixture per pound of
Sensible Heat Factor-The
- Located at 80 F db and 50% rh and used in
conjunction with the sensible heat factor to plot the various
air conditioning process lines.
Alignment Circle
condensa-
of Dry Air-The basis for all psychrometric calculations.
Remains constant during all psychrometric processes.
Pounds
- Ratio of the actual water vapor pressure of
the air to the saturated water vqpor pressure of the air at the
same
temperature.
Relative Humidity
The dry-bulb, wet-bulb, and clewpoint temperatures and the
relative humidity are so related that, if two properties are
known, all other properties shown may then be determined.
When air is saturated, dry-bulb, wet-bulb, and dewpoint temperatures
are all equal.
or Moisture Content-The weight of water vapor
in grains or pounds of moisture per pound of dry air.
Specific Humidity
Enthalpy - A thermal property indicating the quantity of heat
in the air above an arbitrary datum, in Btu per pound of dry
air. The datum for dry air is 0°F and, for the moisture content, 32 F water.
- Enthalpy indicated above, for any given
condition, is the enthalpy of saturation. It should be corrected by the enthalpy deviation due to the air not
being in the saturated state. Enthalpy deviation is in
Btu per pound of dry air. Enthalpy deviation is
qplied where extreme accuracy is required; how. er, on normal air conditioning estimates
it is omitted.
Entholpy Deviation
Dry-Bulb
FIG.
32
ratio of sensible to total heat.
-SKELETON
Temperature
PSYCHROMETRIC
CHART
PSYCHROMETRIC CHART
Normal
Temperatures
AIR CONDITIONING PROCESS
I. RETURN AIR FROM THE ROOM @ IS MIXED WITH
OUTDOOR
2
THIS
AIR
M,XT”RE
@
REOUIRED
OF
OUTDOOR
FOR
AND
VENTILATION.
RETURN
AIR
ENTERS THE APPARATUS @ WHERE IT IS
CONDITIONED
3.
T,,EN
THE
TO
A,R
@
CYCLE
AND
IS
Flc.
SUPPLIED
TO
REPEATED
AGAIN.
33
THE
SPACE
- TYI~ICAL
0.
AIR CONDITIONING P ROCESS T RACED
ON A
S TANDARD PSYCHROMETRIC
.
CHART
l-117
CH,\I’-I‘EK 8 . ,\I’I’LIED I’SYC:HKO~ll:.~T’liI(:S
DEFINITION
Psychrometrics is the science involving thermodynamic properties of moist air and the effect of
atmospheric moisture on materials and human comlort.
it applies to this chapter, the definition must
be broadened to include the method ol controlling
the thermal properties of moist air.
AIR CONDITIONING PROCESSES’
Fig. 33 shows a typical air conditioning process
traced on a psychrometric chart. Outdoor air (2)* is
mixed with return air from the room (I) and enters
the apparatus (3). Air flows through the conditioning apparatus (3 - 4) and is supplied to the space (4).
The air supplied to the space moves along line (4 - 1)
as it picks up the room Ioads, and the cycle is re-
peated. Normally most o[ the air supplied to the
space by the air conditioning system is returned
to the conditioning apparatus. There, it is lnixetl
with outdoor air required Lor ventilation. The mixture then passes tliru tile apparatus where heat and
moisture are added or removed, as required, to
maintain the desired conditions.
The selection of proper equipment
to accomplish
this conclitioning and to control the thcrmotlynanlic
p r o p e r t i e s ot the air depends upon a variety ofelements. However, only those which affect the psychromctric properties of air will bc discussed in this
chapter. These elements are: room sensible heat
factor (RSHFj’)t , grand sensible heat factor (GSHF),
effective surface temperature (tCJ, bypass factor (UF),
and effective sensible heat factor (UHF).
DESCRIPTION OF TERMS, PROCESSES AND FACTORS
SENSIBLE HEAT FACTOR
The thermal properties of air can be separated
into latent and sensible heat. The term sensible
heat factor is the ratio of sensible to total heat, where
total heat is the sum of sensible and latent heat.
This ratio may be expressed as:
SHF=
where: SHF
SH
LH
TH
R’
=
=
=
=
SH
SH
=SHfLH T
H
sensible heat factor
sensible heat
latent heat
total heat
Fig. 34. This line represents the psychrometric process of the supply air within the conditioned space
and is called the room sensible heat factor line.
The slope of the RSHF line illustrates the ratio
of sensible to latent loads within the space and is
illustrated in Fig. 34 by ~h,~ (sensible heat) and AA,
(latent heat). Thus, if adequate air is supplied to
offset these room loads, the room requirements will
M SENSIBLE HEAT FACTOR (RSHF)
The room sensible heat factor is the ratio of room
sensible heat to the summation of room sensible and
room latent heat. This ratio is expressed in the following formula:
RSHF =
RSH
RSH
RSH+RLH=RTH
The supply air to a conditioned space must have
the capacity to offset simultaneously both the room
sensible and room latent heat loads. The room
and the supply air conditions to the space may be
plotted on the standard psychrometric chart and
these points connected with a straight line (1 - 2),
*One italic numljer in parentheses represents a point, and two
italic numl,ers in parentheses represent a line, plotted on the
accompanying psychrometric chart examples.
,
DRY-BULB
TEMPERATURE
/
FIG. M - RSHF LINE P LOTTED BETWEEN ROOM
:$
AND
S UPPLY A IR CONDITIONS
tRefer to page 119 for a tlescription of all al)lneviations
:yml>ols tlsecl in this chapter.
ant1
l-118,
l’.\l<‘I‘ I . LO,\11 ES~I‘IM.\‘I‘IN~;
L
1x2 satislicd, !)rovitl(Yl I)otll tllc dry- and wet-!)ulh
temperatures ol the supply air fall on this line.
T h e ro0111 scnsiblc heat Iactor line cai1 a l s o bc
drawn 011 the psychrometric chart without knowing
the condition ol supply air. The lollowing procetlure illustrates how to plot this line, using the cnlculated RSHF, the room clesign contlitions, the scnsible heat factor scale in the upper right ham! corner
of the psychrometric chart, and the alignment circle
at 80 F dry-!)ulb and 5070 relative humidity:
1. Draw a base line thru the alignment circle and
the calculated RSHF s1~ow11 on the sensible
heat factor scale in the upper right corner of
psychromctric chart (I - _3J, rig. 35.
2. Draw the actual room sensible heat factor line
thru the room design conditions parallel to the
base line in .Step 1 (3 - -/), Fig. 35. As shown, this
line may be drawn to the saturation line 011 the
psychrometric chart.
of the air cnterillg the apparatus (mixture condition
ol outdoor and return room air) ant! the contlition
of the air leaving tlie apparatus may !,e plotted on
the psychrometric chart anr! connected by a straight
litic (/ 21, /‘ix. 36. T h i s line rcl)rcscnts tile psycllt-ometric process ol the air as it passes through the
conditioning apparatus, ant! is relerred to as the
grand scnsi!)lc heat Iactor line.
‘I-lie slope 0C the GSI-11; l i n e represents tlic ratio
oC sensible anrl latent heat that the apparatus must
hantlle. This is illustrated in Pig. 36 !)y A//,? (sensible
heat) ;1nc1 Ah, (Iatcnt lieatj.
/
DESIGN
FROM APPARTUS
DRY-BULB
TEMPERATURE
.
CALCULATEC
RSHF
FIG.
36 - GSHF LINE PLOTTED BETWEEN MIXTURE
CONI~ITIONS T O A PPARATUS AND L E A V I N G
C O N D I T I O N F R O M A PPARATUS
I
/
SOFdb
FIG,
35 - RSHF
LINE PLOTTED
PSYCHROMETRIC
ON
S KELETON
CHART
GRAND SENSIBLE HEAT FACTOR (GSHF)
The grand sensible heat factor is the ratio of the
total sensible heat to the grand total heat load
that the conditioning apparatus must handle, including the outdoor air heat loads. This ratio is
determined from the following equation:
GSHF =
TSH
TSH
T L H + T S H =GTH
Air passing thru the conditioning apparatus
increases or decreases in temperature and/or moisture
content. The amount of rise or fall is determined
by the total sensible and latent heat loads that the
conditioning apparatus must handle. The condition
The grand sensible heat factor line can be plotted
on the psychrometric chart without knowing the
condition of supply air, in much the same manner
as the RSHF line. Fig. 3i, Step I (I - 2) and Step 2
(3 -4) show the procedure, using the calculated
GSHF, the mixture condition of air to the apparatus, the sensible heat factor scale, and’the alighment
circle on the psychrometric chart. The resulting
GSHF line is plotted thru the mixture conditions
of the air to the apparatus.
REQUIRED AIR QUANTITY
The air quantity required to offset simultaneously
the room sensible and latent loads and the air quantity required thru the apparatus to handle the total
sensible and latent loads may be calculated, using
the conditions on their respective RSHF and GSHF
lines. For a particular application, when both the
RSHF and GSHF ratio lines are plotted 011 the psychrometric chart, the intersection of the two lines (I)
Fig. 38, rcprescnts the condition ol‘ the supply air to
CHAPTEK
l-119
8 . ,\I’I’t,IEI> I’SY(:HIIOIVIE’I‘I~I(:S
these supplementary loads are considered in plotting
the RSHF and GSHF lines.
Point (I) is the condition of air lcaving the apparatus and point (2) is the condition of supply air
to the space. Line (1 - 2) represents the temperature
rise of the air stream resulting from fan horsepower
and heat gain to the duct.
OUTDOOR
DESIGN
/
DRY-BULB
r
IG.
I
MIXTURE
/
CONDITION TO
APPARATUSt, I c
is
I
1
BOF
TEMPERATURE
37 - GSHF L INE P LOTTED
PSYCHROMETRIC
ON
A-1CONDlTlON OF
OF AIR
AIR LEAVING
LEAVING APPARATUSt
APPARATUSt
S KELETON
2
8
%
f/&,1
CH A R T
DRY-BULB
1
I
TEMPERATURE
*
OUTDOOR
FIG.
39 - RSHF
WITH
AND
GSHF L INES P LOTTED
S UPPLEMENTARY
L OAD
L INE
The air quantity required to satisfy the room load
may be calculated from the following equation:
cfm,, =
ROOM AND AIR
LEAVING APPARATUS
DRY-BULB
FIG.
38 - RSHF
S KELETON
AND
I
TEMPERATURE
GSHF L INES P LOTTED
P SYCHROMETRIC
RSH
1.08 (Ln - LJ
The air quantity required thru the conditioning
apparatus to satisfy the total air conditioning load
(including the supplementary loads) is calculated
from the following equation:
ON
CHART
the space. It is also the condition of the air leaving
the apparatus.
This neglects fan and duct heat gain, duct leakage
losses, etc. In actual practice, these heat gains and
losses are taken into account in estimating the cooling load. Chapter 7 gives the necessary data for evaluating these supplementary loads. Therefore, the
temperature of the air leaving the apparatus is not
necessarily equal to the temperature of the air supplied to the space as indicated in Fig. 38.
Fig. 39 illustrates what actually happens when
Cfmda
=
TSH
1.08 (L - trd
The required air quantity supplied to the space
is equal to the air quantity required thru the apparatus, neglecting leakage losses. The above equation contains the term t,,, which is the mixture
condition of air entering the apparatus. With the
exception of an all outdoor air application, the
term t, can only be determined by trial and error.
One possible procedure to determine the mixture
temperature and the air quantities is outlined below.
This procedure illustrates one method of apparatus
selection and is presented to show how cumbersome
and time consuming it may be.
A
l-120
1.
I’:\RT I . LOAD
a rise (trm - t,,J in the supply air to the
and calculate the supply air quantity
(cfm,,) to the space.
Assume
SpaCC,
2. Use this air quantity to calculate the mixture
condition of the air (t,,,) to the space, (Equation
1, p~lge 150).
3. Substitute this supply air quantity and mixture
condition of the air in the formula for air
quantity thru the apparatus (cfm,,) and determine the leaving condition of the air from the
conditioning apparatus (t,,,).
4. The rise between the leaving condition from
the apparatus and supply air condition to the
space (L - tl,,) must be able to handle the
supplementary loads (duct heat gain and fan
heat). These temperatures (t,,,, t,,) may be
plotted on their respective GSHF and RSHF
lines (Fig. 39) to determine if these conditions
can handle the supplementary loads. If they
cannot, a new rise in supply air is assumed and
the trial-and-error procedure repeated.
In a normal, well designed, tight system this difference in supply air temperature and the condition
of the air leaving the apparatus (t,, - t,& is
usually not more than a few degrees. To simplify
the discussion on the interrelationship of RSHF and
GSHF, the supplementary loads have been neglected
in the various discussions, formulas and problems
in the remainder of this chapter. It can not be overemphasized, however, that these supplementary
loads must be recognized when estimating the cooling and heating loads. These loads are taken into
account on the air conditioning load estimate in
Chapter 1, and are evaluated in Chapter 7.
The RSHF ratio will be constant (at full load)
under a specified set of conditions; however, the
GSHF ratio may increase or decrease as the outdoor
air quantity and mixture conditions are varied for
design purposes. As the GSHF ratio changes, the
supply air condition to the space varies along the
RSHF line (Fig. 38).
The difference in temperature between the room
and the air supply to the room determines the air
quantity required to satisfy the room sensible and
room latent loads. As this temperature difference
increases (surlplying colder air, since the room conditions are fixed), the required air quantity to the
space decreases. This temperature difference can
increase up to a limit where the RSHF line crosses
the saturation line on the psychrometric chart, Fig.
38; assuming, of course, that the available conditioning equipment is able to take the air to 100%
ES’I’IMATING
saturation. Since this is impossible, the condition of
the air normally falls on the RSHF line close to
the saturation line. How close to the saturation line
depends on the physical operating characteristics
and the efficiency of the conditioning equipment.
In determining the required air quantity, when
neglecting the supplementary loads, the supply air
temperature is assumed to equal the condition of the
air leaving the apparatus (tY,l - tldb). This is illustrated in Fig. 38. The calculation for the required
air quantity still remains a trial-and-error procedure, since the mixture temperature of the air
(t,,,) entering the apparatus is dependent on the
required air quantity. The same procedure previously described for determining the air quantity is
used. Assume a supply air rise and calculate the
supply air quantity and the mixture temperature to
the conditioning apparatus. Substitute the supply l
air quantity and mixture temperature in the equation for determining the air quantity thru the
apparatus, and calculate the leaving condition of
the air frbm the apparatus. This temperature must
equal the supply air temperature; if it does not, a
new supply air rise is assumed and the procedure
r e p e a t e d .
Determining the required air quantity by either
method previously described is a tedioui process,
since it involves a trial-and-error procedure, plotting
the RSHI; and GSHF ratios on a pspchrometric
chart, and in actual practice accounting for the
supplementary loads in determining the supply air,
mixture and leaving air temperatures.
This procedure has been simplified, however, by
relating all the conditioning loads to the physical
performance of the conditioning equipment, and
then including this equipment performance in the
actual calculation of the load.
This relationship is generally recognized as a
psychrometric correlation of loads to equipment performance. The corrtilation is accomplished by calculating the “effective surface temperature,” “bypass
factor” and “effective sensible heat factor.” These
alone will permit the simplified calculation of stiipply air quantity.
EFFECTIVE SURFACE TEMPERATURE (fJ
The surface temperature of the conditioning
equipment varies throughout the surface of the apparatus as the air comes in contact with it. However,
the effective surface temperature can be considered
to be the uniform surface temperature which would
produce the same leaving air conditions as the nonuniform surface temperature that actually occurs
when the apparatus is in operation. This is more
clearly understood by illustrating the heat transfer
effect between the air and the cooling (or heating)
medium. Fig. 40 illustrates this process and is applicable to a chilled water cooling medium with the
supply air counterflow in relation to the chilled
dewpoint (adp). The term is used exclusively in this
chapter when relcrring to cooling and dchumidifying applications. The psychrometrics of air can be
applied equally well to other types of heat transfer
applications such as sensible heating, evaporative
cooling, sensible cooling, etc., but for these applications the effective surface temperature will not
necessarily fall on the saturation line.
BYPASS FACTOR (BF)
Bypass factor is a function of the physical and
operating characteristics of the conditioning apparatus and, as such, represents that portion of the air
which is considered to pass through the conditioning
apparatus completely unaltered.
SURFACE
AREA
FIG. 40 - RELATIONSHIP OF E FFECTIVE SURFACE TEMP
TO SUPPLY A IR AND CHILLED W ATER
The relationship shown in Fig. 40 may also be
illustrated for heating, direct. expansion cooling and
for air *flowing parallel to the cooling or heating
medium. The direction, slope and position of th,e
lines change, but the theory is identical.
Since conditioning the air thru the apparatus reduces to the basic principle of heat transfer between
the heating or cooling media of the conditioning
apparatus and the air thru that apparatus, there
must be a common referen$e point. This point is
the effective surface temperature of the apparatus.
The two heat traiisfers are relatively independent of
e2-h other, but are quantitatively equal when re1.
d to the effective surface temperature.
Therefore, to obtain the most ‘economical apparatus selection, the effective surface temperature is
used in calculating the required air quantity and in
selecting the apparatus.
For applications involving cooling and dehumidification, the effective surface temperature is at the
point where the GSHF line crosses the saturation
line on the psychrometric chart (Fig. 36). As such,
this effective surface temperature is considered to be
the dewpoint of the apparatus, and hence the term
apparatus dewpoint (adp) has come into common
usage for cooling and dehumidifying processes.
Since cooling and dehumidification is one of the
most common applications for central station apparatus, the “Air Conditiouing Load Estimate” form,
Fig. 44, is designed around the term apparatus
The physical and operating characteristics affecting the bypass factor are as follows:
1. A decreasing amount of available apparatus
heat transfer surface results in an increase in
bypass factor, i.e. less rows of coil, less coil
surface area, wider spacing of coil tubes.
2. A decrease in the velocity of air through the I
conditioning apparatus results in a decrease
in bypass factor, i.e. more time for the air to
contact the heat transfer surface.
Decreasing or increasing the amount of heat transfer surface has a greater effect on bypass factor than
varying the velocity of air through the apparatus.
There is a psychrometric relationship of bypass
factor to GSHF and RS$IF. Under specified room,
outdoor design conditions and quantity of outdoor
air, RSHF and GSHF are fixed. The position of
RSHF is also fixed, but the relative position of
GSHF may vary as the supply air quantity and
supply air condition change.
To properly maintain room design conditions,
the air must be supplied to the space at some point
along the RSHF lihe. Therefore, as the bypass factor
varies, the relative position of GSHF to RSHF
changes, as shown by the dotted lines in Fig. 41. As
the position of GSHF changes, the entering and
leaving air conditions at the apparatus, the required
air quantity, bypass factor and the apparatus dewpoint also change.
The effect of varying the bypass factor on the
conditioning equipment is as follows:
1. Smaller bypass factor a. Higher adp - DX equipment selected for
higher refrigerant temperature and chilled
water equipment would be selected for less b&Y
or higher temperature chilled water. Possibly smaller refrigeration machine.
l-122
l’,jRT
1,. Less air - smaller fan and fan motor.
c. More heat transfer surface - more rows of
coil or more coil surface available.
d. Smaller piping if less chilled water is used.
2. Larger bypass factor a. Lower adp - Lower refrigerant temperature
to select DX equipment, and more water or
lower temperature for chilled water equipment. Possibly larger refrigeration machine.
b. Marc air -larger fan and fan motor.
c. Less heat transfer surface - less rows of coil
or less coil surface available.
d. Larger piping if more chilled water is used.
I. LOAD ESTIMATING
rically to the bypass factor. Although it is recognized
c.hat bypass factor is not a true straight line function,
it can be accurately evaluated mathematically from
the following equations:
BF=
’ Idb
- hdp
t edb - &do
fl,a
=
- hndp
he,
wla - Wadp
- hadp
=
$,a - wad,
and
NOTE:
The quantity (I--UF) is frequently called contact factor
and is considered to be that portion of the air leaving
the apparatus at the adp.
EFFECTIVE SENSIBLE HEAT FACTOR (ESHF)
OUTDOOR
DESIGN
/
AIR
DRY-BULB
FIG .
To relate bypass factor and apparatus dewpoint
to the load calculation, the eflective sensible heat
factor term was developed. ESHF is interwoven with
BF and adp, and thus greatly simplifies the calculation of air quantity and apparatus selection.
The effective sensible heat factor is the ratio of
effective room sensible heat to the effective room
sensible and latent heats. Effective room sensible
heat is composed of room~sensible heat (seePSHF)
plus that portion of the outdoor air<&ible load
which is considered a’s being bypassed, unaltered,
thru the conditioning apparatus. The effective room
latent heat is composed of the room latent heat
(see RSHF) plus that portion of the outdoor air
latent heat load which is considered as being bypassed, unaltered, thru the conditioning apparatus.
This ratio is expressed in the following formula:
LEAVING
TEMPERATURE
41 - RSHF AND GSHF LINES P LOTTED
SKELETON PSYCHROMETRICCHART
ON
It is, therefore, an economic balance of first cost
’ and operating cost in selecting the proper bypass
factor for a particular application. Table 62, page
127, lists suggested bypass factors for various applications and is a guide for the engineer to proper bypass
factor selection for use in load calculations.
Tables have also been prepared to illustrate the
various configurations of heat transfer surfaces and
the resulting bypass factor for different air velocities.
Table 61, pnge 127, lists bypass factors for various
coil surfaces. Spray washer equipment is normally
rated in terms of saturation efficiency which is the
complement of bypass factor (1 - BF). Table 63,
page 136, is a guide to representative saturation efficiencies for various spray arrangements.
As previously indicated, the entering and leaving
air conditions at the conditioning apparatus and
the apparatus dewpoint are related psychromet.
/
ESHF =
ERSH
ERSH + ERLH
ERSH
=ERTH
The bypassed outdoor air loads that are included
in the calculation of ESHF are, in effect, loads imposed on the conditioned space in exactly the same
manner as the infiltration load. The infiltration
load comes thru the doors- and windows; the bypassed outdoor air load is supplied to the space
thru the air distribution system.
Plotting RSHF and GSHF on the psychrometric
chart defines the adp and BF as explained previously. Drawing a straight line between the adp and
room design conditions (1 - 2), Fig. 42 represents the
ESHF ratio. The interrelationship of RSHF and
GSHF to BF, adp and ESHF is graphically illustrated in Fig. 42.
The effective sensible heat factor line may also
be drawn on the psychrometric chart without initialIy knowing the adp. The procedure is identical
to the one described for RSHF on @ge IIN. The cal-
l
1-123
(:H/\I'TEK 8. .\l'l'Ll131) I'SYC:l-IROILIE'I‘I~I<:S
c&ted ESHF, however, is plotted thru the room
design conditions to the saturation line (I - 2), I;ig.
43, thus indicating the adp.
Tables have been prepared to simplify the method
of determining adp from ESHF. Adp can be obtained by entering Table 65 at room design conditions and at the calculated ESHF. It is not necessary
to plot ESHF on a psychrometric chart.
AIR QUANTITY USING ESHF, ADP AND BF
A simplified approach for determining the required air quantity is to use the psychrometric corre-
OUTDOOR
lation of effective sensible heat factor, apparatus
dewpoint and bypass factor. Previously in this chapter, the interrelationship of ESHF, BF and adp was
shown with GSHF and RSHF. These two factors
need not be calculated to determine the required
air quantity, since the use of ESHF, BF and adp
results in the same air quantity.
The formula for calculating air quantity, using
BF and tndp, is:
Cf%a =
ERSH
1.08 (trm - tad (1 - I-3
(ESHF is used to determine tadp.)
This air quantity simultaneously offsets the room
sensible and room latent loads, and also handles the
total sensible and latent loads for which the conditioning apparatus is designed, including the outdoor
air loads and the supplementary loads.
AIR CONDITIONING LOAD ESTIMATE FORM
FIG. 42 - RSHF, GSHF AND ESHF LINES PLOTTED
ONSKELETON
PSYCHROMETRIC
CHART
The “Air Conditioning Load Estimate” form
is designed for cooling and dehumidifying applications, and may be used for psychrometric calculations. Normally, only ESHF, BF and adp are
required to determine air quantity and to select
the apparatus. But for those instances when it is
desirable to know RSHF and GSHF, this form is
designed so that these factors may also be calculated.
Fig. 44, in conjunction with the following items,
explains how each factor is calculated. (The circled
numbers correspond to numbers in Fig. 44.) ._
1. RSHF =
RSH
0
=0+0
RSH + RLH
2. G S H F = = = @+O
GTH
3. ESHF =
0
ERSH
ERSH.+ E R L H
ERSH
=ERTH
4. Adp located where ESHF crosses the saturation
line, or from Table 65. ESHF @ and room
conditions @ give adp @.
SOF
DRY-BULB
FIG.
TEMPERATURE
43-ESHF LINEPLOTTEDONSKELETON
PSYCHROMETRICCHART
5. BF $J used in the outdoor air calculations is
obtained from the equipment performance
table or charts. Typical bypass factors for different surfaces and for various applications are
given on page 127. These are to guide the engineer and may be used in the outdoor air
calculation when the actual equipment per- . ’
formance tables are not readily available.
I
1-124
P.411'I‘ 1. LO,\11 ES'I‘IM.\'I‘ING
i
DATE----~.
SHEET
P R E P A R E D
NAME OF
-yOFFlCE
BY--
PROP NO
B
------JOaNO.
JOB
L O C A T I O N - - - SPACE
USED
FOR
S o FT %
..-‘LISS
GLASS
_ So FT
GLASS
SKVLIGHT
X
So FT x
WALL
s a FT x
x
WALL
Sa FT y
x
WALL
SQ FT x
x
Sa FT X
x
Sa FT X
x
PLOPLE
JLL G LASS
S o FT x
x
JmITION
Sa FT x
x
_CgLING
Sa
FLOOR
SP FT x
Fr
-~
CFM
OUTDOOR
AIR
EFFECTlYE
SENS HEIT =
FACTOR
._
HPoaKW X
HE/T
SELECTED ADP =
AIR
.._
-.. F
QUANTITY
EFFECTIYE
R O O M SENS.
H E A T __v.,w-~CFM~~
____1.08
x
-
-
0
- -
-
W A T T S x 1.4 Y
LlGHTS
-..-.---.-CFMo~
0
(1 --RF)
X tT,,-&F - T,,,v.@vF) = -F
8
POWER
..- F
DEHUMIDIFIED
INTERNAL HEAT
PEOPLE
q
APPARATUS
EFFECTIVE R O O M T O T A L H E A T = .--
0
INDICATED ADP =
1.08
x
THRU
APPARATUS DEWPOINT
EFFECT,YE
O O M SENS.
8
--.-- R _-...
.-
-x
PEOPLE
ADDITION&L
= --_
-
x
x
cm x
APPLIANCES.
lxxx!
.-x.--
TRANS. GAIN-EXCEPT WALLS b ROOF
INFlLTRATlON
lxxx
OUTDOOR AIR
x -CFY,PLIJOH
iv ROOF
WALL
ROOF-SHADED
lxxx
-.X
Y
s o FT Y
&_OF-SUN
I
x
SOLAR & TRANS. GAIN-WALLS
-
YCE
-x-
SC) F T x
-
x
@
F
TLMP
~1st
ROOW SENT. H E A T
- ~-= --F~I+,UTLETLII~,.
x
@
CFM D.
-
1.08
x
ETC .
SUPPLY AIR QUANTITY
x
HE A T G A I N S
R O O M SCNS. H E A T
-
-52
SUB TOTA-
1.08
x
F
DESIRED
-
-
=--CFM5,
~,r,
43
8
PCFM
5* - -----CFMDA
=-.-CFMo*
EDa
LDS
RESULTING ENT t LVG CONDITIONS AT APPARATUS
@ CFM,,&
@ - T,,-F)
0
f------ X CT,,-F
= T,,,-F
0
T,,-F
@oO.@cFM+
@
0
T,,--.-Ff-aF
FROhi
LATENT HEAT
CFM
X
lNF!LTRATlON
PEOPLE
PlOPLE
STEAM
LB,“”
APPLIAHCES.
Aoo,r,ow~~
G”,Lrn
CHART:
x 0.m
@ X (T,D,-F
TAD-F)
=
T,,,-F
T,,.- F. T LWF----F
NOTES
x
x 1030
E TC .
HE A T G A I N S
so FT x lilO0
VAPOR TRINS.
x
GIiLB x
___-- SUB T
SAFETY
PSYCH.
F ACTOR
O T A L
%
ROOM LATENT HEAT
S U P P L Y
Ourooo~
D U C T
L E A K A G E
EFFECTlVE
E F F E C T I V E
SENSIBLE:
LATENT:
RLT”l”
DUCT
“CAT Gl,H
L E . - p - %
GR,LB
CFH x
AIR
x
0 BF
x
0.68
R O O M
T O T A L
H E A T
n
OUTDOOR AIR HEAT
Fx(,---@aF)Xl.OS.
CFM x
CFM
x
RET”“”
DUCT
% + LLLS. Gil”
p
,@
ROOM LATENT HEAT
G”,UX (,- II
HP
?/. + PUMP
G R A N D
BF)
/
/@
,-l-x
Y 0.6&-AL;k+y)
SUB
T OTAL
DC”““. *
a/& + PlPL LOIS
%./@
T O T A L
H E A T
I
Form E 20
NOTE: The circled numbers are explained on the previous page under “Air Conditioning Load Estimate”
FIG. 44 -AAIR~ONDITIONING
LOAD ESTIMATE
form.
,
l-125
CHAI’~I‘IIlI 8 . .\I’I’LIED I’SY(;~lIIO~lE’I‘KI(:S
leaving air conditions are easily determined.
The calculations for the entering and leaving
dry-bulb temperatures at the apparatus are
illustrated in Fig. 44.
ERSH
6. cfm,,,= ___
1.08 (r,, - cd (1 - B6
The entering dry-bulb calculation contains the
This air quantity “cfmt” determ “cfmt”‘.
pends on whether a mixture of outdoor and
return air or return air only is bypassed around
the conditioning apparatus.
( Once the dehumidified air quantity is calcu\ lated, the conditioning apparatus may be se’ lected. The usual procedure is to use the grand
total heat @ , dehumidified air quantity
:a, and the apparatus dewpoint @$ , to
4 select the apparatus.
The total supply air quantity cfm,q, @ is used
for “cfmt” when bypassing a mixture of outdoor and return air. Fig. 45 is a schematic sketch
of a system bypassing a mixture of outdoor
and return air.
Since guides are available, the bypass factor
of the apparatus selected is usually in close
agreement with the originally assumed bypass
factor. If, because of some peculiarity in loading in a particular application, there is a wide
divergence in bypass factor, that portion of the
load estimate form involving bypass factor
should be adjusted accordingly.
7. Outlet temperature difference - Fig. 44 shows
a calculation for determining the temperature
difference between room design dry-bulb and
the supply air dry-bulb to the room. Frequently
a maximum temperature difference is established for the application involved. If the outlet
temperature difference calculation is larger
than desired, the total air quantity in the
system is increased by bypassing air around
the conditioning apparatus. This temperature
difference calculation is:
Outlet temp diff =
=
8.
CONDlTlONEO
(
BYPASSING
-
=
The amount of air that must be bypassed
around the conditioning apparatus to maintain
this desired temperature difference (At) is the
difference between cfm,, and cfmda.
9. Entering and leaving conditions at the apparatus - Often it is desired to specify the selected
conditioning apparatus in terms of entering
and leaving air conditions at the apparatus.
Once the apparatus has been selected from
ESHF, adp, BF and GTH, the entering and
f
FAN
ENT
CONOITIONING
COND
FIG .
45 - BYPASSING MIXTURE
RETURN
OF
OUTDOOR
AND
AIR
When bypassing a mixture of return air only
or when there is no need for a bypass around
the apparatus, use the cfmda @ for the value
of “cfmt”. Fig. 46 is a schematic sketch of a system bypassing room return air only.
0
x @
RSH
“\
@
1.08 X At ?’ 1.08 X At
c
1
Total air quantity when outlet temperature
difference is greater than desired - The calculation for the total supply air quantity for a
desired temperature difference (between room
and outlet) is:
cfm,,
.
i.1 ’
t
RSH
1.08 x cfmda
1.08
SUPPLY
AIR
MIXTURE
OF OUTDOOR AN0
RETURN AIR
CONDlTlONED
SPACE
.
SUPPLY
AIR
BYPASSING
RETURN
AIR
-
t
FAN
OUTDOOR
NR
-
FIG .
ENT
CON0
CONDITIONING
APPARATUS
46 - BYPASSING RETURN A IR ONLY
OR
No FIXED BYPASS
*“cfmt” is a symbol appearing in the equation next to 0 in
Fig. 44.
I
’
1-126
I’AK’I I. LOAD ESTIMATING
this point. (This point delincs the intcrsection of the RSHF and GSHF as described
previously.)
The entering and leaving wet-bulb temperatures at the apparatus are determined on the
standard psychrometric chart, once the entering and leaving dry-bulb temperatures are calculated. The procedure for determining the
wet-bulb temperatures at the apparatus is illustrated in Fig. 47 and described in the following
i terns:
a. Draw a straight line connecting room design
conditions and outdoor design conditions.
b. The point at which entering dry-bulbcrosses
the line plotted in Step a defines the entering conditions to the apparatus. The entering wet-bulb is read on the psychrometric
chart.
c. Draw a straight line from the adp @ to
the entering mixture conditions at the apparatus (Step IJ.) (This line defines the GSHF
line of the apparatus.)
d. The point at which the leaving dry-bulb
crosses the line drawn in Step c defines the
leaving conditions of the apparatus. Read
the leaving wet-bulb from the apparatus at
CALCULATED
LEAWNG DRY BULB TEMP
CALCULATED
ENTERING DRY BULB TEMP
FIG. 47 - ENTERING
AT
AND L EAVING CONDITIONS
APPARATUS
AIR CONDITIONING APPARATUS
The following section describes the characteristic
pspchrometric performance of air conditioning
equipment.
Coils; sprays and sorbent dehumidifiers are the
three basic types of heat transfer equipment required for air conditioning applications. These
components may be used singly or in combination
to control the psychrometric properties of the air
passing thru them.
The selection of this equipment is normally determined by the requirements of the specific application. The components must be selected and
integrated to result in a practical system; that is,
one having the most economical owning and operating cost.
An economical system requires the optimum combination of air conditioning components. It also
requires an air distribution system that provides
good air distribution within the conditioned space,
using a practical rise between supply air and room
air temperatures.
Since the only known items are the load in the
space and the conditions to be maintained within
.
the space, the selection of the various components
is based on these items. Normally, performance requirements are established and then equipment is
selected to meet the requirements.
COIL
CHARACTERISTICS
In the operation of coils, air is drawn or forced
over a series of tubes thru which chilled water,
brine, volatile refrigerant, hot water or steam is flowing. As the air passes over the surface of the coil, it
is cooled, cooled and dehumidified, or heated, depending upon the temperature of the media flowing
thru the tubes. The media in turn is heated or
cooled in the process.
The amount of coil surface not only affects the
heat transfer but also the bypass factor of the
coil. The bypass factor, as previously explained, is
the measure of air side performance. Consequently,
it is a function of the type and amount of coil
surface and the time available for contact as the
air passes thru the coil. Table 61 gives approximate
bypass factors for various finned coil surfaces and
air velocities.
CHAI'TEK
1-127
8 . /~I'PI,IEI) I'SYCIiROMETKIC:S
TABLE 6 1 -TYPICAL BYPASS FACTORS
(For Finned Coils)
DEPTH
OF
COILS
WITHOUT
SPRAYS
SPRAYS;
8 lins/in.
14 fins/in.
300-700
300-700
8 fins/in. 14 fins/in.
Velocity
(rows)
WITH
(fpm)
300 -700
300 - 700
3
.42 - .55
.21 * .40
.22 - .38
.I0 - .23
4
.I9 - .30
.05 - .14
.12 - .23
.02 - .09
.I2 - .22
.os - .14
.03 - .lO
5
6
.08 - .18
.Ol - .06
.06 - .ll
.Ol - ;05
8
.03 - .08
2
.Ol - .08
.02 - .05
*The bypass factor with spray-coils is decreased because
spray provides more surface for contacting the air.
the
DRY-BULB
These bypass factors apply to coils with s/s in.
c
. tubes and spaced on approximately 11/d in.
centers. The values are approximate. Bypass factors
for coils with plate fins, or for combinations other
than those shown, should be obtained from the coil
manufacturer.
Table 61 contains bypass factors for a wide range
of coils. This range is offered to provide sufficient
latitude in selecting coils for the most economical
system. Table 62 lists some of the more common
applications with representative coil bypass factors.
This table is intended only as a guide for the design
engineer.
TABLE 62-TYPICAL BYPASS FACTORS
(For
-
COIL
BYPASS
- ’ CTOR
Various
Applications)
TYPE OF APPLICATION
EXAMPLE
FIG . 48 - COIL P ROCESSES
COIL
PROCESSES
Coils are capable of heating or cooling air at a ,
constant moisture content, or simultaneously cooling and dehumidifying the air. They are used to
control dry-bulb temperature and maximum relative humidity at peak load conditions. Since coils
alone cannot raise the moisture content of the air,
a water spray on the coil surface must be added
if humidification is required. If this spray water is
recirculated, it will not materially affect the psychrometric process when the air is being cooled and
dehumidified.
Fig. 48 illustrates the various processes that can be
accomplished by using coils.
Sensible
0.30 to 0.50
A small total load or a load
that is somewhat larger with
a low sensible heat factor
(high latent load).
0.20 to 0.30
Typical comfort application ~
Residence,
with a relatively small total
Small
load or a low sensible heat
factor with a somewhat larger Retail Shop,
Factory
load.
0.10 to 0.20
Typical
0.05 to 0.10
Applications with high internal sensible loads or requiring
a large amount of outdoor air
for ventilation.
0 to 0.10
All
comfort
outdoor
air
application.
applications.
I
I I<esidence
c
I
Dept. Store,
Bank, Factory
Dept. Store,
Restaurant,
Factory
Hospital
Operating
Room, Factory
TEMPERATURE
Cooling
The first process, illustrated by line (1 - Z), represents a sensible cooling application in which the
heat is removed from the air at a constant moisture
content.
Cooling
and
Dehumidification
Line (1 - 3) represents a cooling and dehumidification process in which there is a simultaneous
removal of heat and moisture from the air.
For practical considerations, line (1 - 3) has been
plotted as a straight line. It is, in effect, a line that
starts at point (I) and curves toward the saturation
line below point (3). This is indicated by line (I - 5).
Sensible
Heating
Sensible heating is illustrated by line (1 - 4); heat
is added to the air at constant moisture content.
PAR?’ 1. LOAD ESTIMATINd
1-128
COIL PROCESS EXAMPLES
To better understand these processes and their
variations, ;1 description of each with illustrated
examples is presented in the following: (Refer to
page 149 for tle finit’ 1011 of symbols and abbreviations.)
Cooling and Dehumidification
Cooling and dehumidification is the simultaneous
removal of the heat and moisture from the air, line
(1 - 3), Fig. 48. Cooling and dehumidification occurs
when the ESHF and GSHF are less than 1.0. The
ESHF for these applications can vary from 0.95,
where the load is predominantly sensible, to 0.45
where the load is predominantly latent.
RSH - 200.000 Btu/hr
RLH - 50,000 Btu/hr
Ventilation - 2,000 cfm,,
Find:
1.
2.
3.
4.
5.
6.
Outdoor air load (OATH)
Grand total heat (GTH)
Effective sensible heat factor (ESHF)
Apparatus dewpoint
temperature (tad,,)
Dehumidified air quantity (cfm,,)
Entering and leaving conditions at the apparatus
(t e d b ’ et w b ' tldb’ hub)
Solution:
1. OASH = 1.08 X 2000 X (95 - 75) = 43200 Btu/hr
OALH = 68 X 2000 X (99 - 65) = 46,200 Btu/hr
OATH = 43,200 -I- 46,200 = 89,400 Btu/hr
(14)
(‘5)
(17)
2. TSH
= 200,000 +,43,200 = 243,200 Btu/hr
TLH = 50,000 + 46,200 = 96,200 Btu/hr
G T H = 243,200 + 96,200 = 339,400 Btu/hr
The air conditioning load estimate form illus&rated in I’ig. 44 presents the procedure that is used
.o determine the ESHF, dehumidified air quantity,
and entering and leaving air conditions at the apparatus. Example 1 illustrates the psychrometrics
involved in establishing these values.
(7)
(8)
(9)
l
3. Assume a bypass factor of 0.15 from Table 62.
200,000 + (.15) (43,200)
ESHF = 200,000 + (.15) (43,200) + 50,000 f (.15) (46,200)
= .785
(26)
4. Determine the apparatus dewpoint
from the room design
conditions and the ESHF, by either plotting on the psychrometric chart or using Table 65. Fig. 49 illustrates the
ESHF plotted on the psychrometric
chart.
Example 1 - Cooling and Dehumidification
Given:
Application - 5P & lO$ Store
Location - Bloomfield, N. J.
t a@
=50F
1
NOTE: Numbers in parentheses at right edge of column refer
to.,equations beginning on page 150.
Summer design - 95 F db, 75 F wb
Inside design - 75 F db, 50% rh
OUTDOOR
75 Fdb
54.4 F db
FIG.~~
--
COOLING
AND
79.45 F db
I
9 S F db
DEHUMIDIFICATION
(I:H/\I”I‘k:K
8 . /\I’l’I,lEl)
l-129
I’SY~:l-lKO~I~‘1‘K1~:S
6. /\sstcmc lor this example that the
9,000 t fm, 50 F adp, and G’rII =
factor that is equal, or nearly equal,
0.15. :\lso, assume that it is not
I~ypass air around the ;ippara”Is,
apl>aratus selected for
339,400, has a I)ypass
lo the assumctl BF =
necessary to physically
‘(
t cd6 -- (2000 x 95) + (7000 x 75)
(31)
,3. I:, I* (Ill
9000
Read tolo where the lcdb crosses tile straigtlt lint plottetl
I)etuvcctl tllc 0ut(l0or and roan, tlcsign contlitions on tllc
psychromctric chart, Fig.
t ewh
I’).
= f35.5 F WI)
Ild6 = 50 + .I5 (79.45 - 50) = 54.4 F tll,
(32)
lktcrminc the tizob
I)y drawing a straight line between
the atlp and the entering conditions at the apparatus.
(This is the GSHF line.) Whcrc lldh intersects this lint,
read tlwb.
t ltuh
Dehumidification
Example
-
High
Latent
Load
cIli some applications a special situation exists
if the ESHF and GSHF lines do not intersect the
saturation line when plotted on the psychrometric
chart or if they do the adp is absurdly low. This
may occur where the latent load is high with respect
to the total loads (dance halls, etc.). In such applications, an appropriate apparatus dewpoint is selected and the air is reheated to the RSHF line.
Occasionally, altering the room design conditions
eliminates the need for reheat, or reduces the quantity of reheat required. Similarly, the utilization of
a large air side surface (low bypass factor) coil may
eliminate the need for reheat or reduce the required
reheat.
Once the ventilation air requirement ,is determined, and if the supply air quantity is not fixed,
the best approach to determining the apparatus
dewnoint is to assume a maximum allowable temPel. .re difference between the supply air and the
room. Then, calculate the supply air conditions to
the space. The supply air conditions to the space
must fall on the RSHF line to properly offset the
sensible and latent loads in the space.
There are four criteria which should be examined, to aid in establishing the supply air requirements to the space. These are:
1. Air movement in the space.
2. Maximum temperature difference between the
supply air and the room.
3. The selected adp should provide an economical refrigeration machine selection.
4. In some cases, the ventilation air quantity
required may result in an all outcloor air
application.
2
-
Cooling
and Dehumidification -
High
L a t e n t load
Givcll:
i\pplication
- Lai)oramry
Location - Bangor, Maine
Summer design - 90 F (II), 73 F wl)
Insitlc &sign - 75 1: (II). .50C,‘:, rh
RSH - 120,000 Iltr+r
RLII - 65,000 Btu/hr
Ventilation - 2,500 cfj?l,,
Temp. tliff. Ixtwccn room and supply air, 20 F maximum
Find:
I. Outdoor air load (O:\TH)
2. Efkctivc scnsiblc heat factor (ESHF)
3.
= 52 . j F WI)
Cooling and
Ar
‘:ation
Extrtnple
_Q is ;I laboratory application with a high
latent load. In this example the ESHF intersects the
saturation line, but the resulting atlp is too low.
/\pparatL!S
LkwpOin~
(t&p)
4. Reheat required
5. Supply air quantity (cfttlsa)
6.
Entering
Conditions
to C o i l
(t&b,
telob,
6ven)
7. Leaving conditions from coil (tldb, tllub)
8. Supply air condition to the space (fsn, 1%)
9. Grand total heat (GTH)
Solution:
1. OASH = I.08 X 2.500 X (90.75) = 40,500 Btu/hr
OALH = .68 X 2500 X (95-65) = 51,000 Btu/hr
Oi\l‘H = 40,500 f51,000 = 91,500 Btu/hr
2. r\ssume
(‘4)
(15)
(17)
a bypass factor of 0.05 because of high latent load.
120,000 + (.05) (40,500)
120,000 + (.05) (40,500) + 65,000 + (.05) (51,000)
= ,645
(26)
ESHF =
When plotted on the psychrometric chart, this ESHF
(.G45 intersects the saturation curve at 35 F. With such
a low adp an appropriate apparatus dewpoint s h o u l d
he selected and the air reheated to the RSHF line.
3. Refer to Table 65. For inside design conditions of 75 F
db, SOSr, rh, an ESHF of .74 results in an adp of 48 F
which is a reasonable minimum figure.
4. Determine amount of reheat (Btu/hr) required to produce an ESHF of .74.
ESHF (74) =
120,000 + .05 (40.500) + reheat
1 2 0 , 0 0 0 + .05 (40,500) + reheat + 65,000 + (.05) 51,000
,74= 122,025 + reheat
(25)
189,575 + reheat
reheat = 70,230 Btu/hr
5. Determine clrhumidilicr
air quantity (‘cf?!lda)
ERSH
cfmda = 1 .OY X (I - BF) (tr,,b - t”,,,,)
122,025 + 70,230
=
= 6940 cfm
1.08 (1 - .05) (75-48)
when no air is to be physically byCfm&
iS ak0 Cfmsa
passed around the cooling coil.
G. t db = (2500 X 90) + (4440 X 75)
e
6940
= 80.4
(31)
SOTE: Numbers in parentheses at right edge of column refel
to equa’ions beginning on fxzge 150.
PART I. L O A D E S T I M A T I N G
l-130
i
Cooling and Dehumidification - Using All Outdoor Air
I n some applications it may be necessary to sup-
pIy a l l o u t d o o r air; for example, a hospital operating room, or an area that requires large
quantities of ventilation air. For such applications,
the ventilation or code requirements may be equal
to, or more than, the air quantity required to handle the room loads.
Items I thrz~ 5 explain the procedure for determining the dehumidified air requirements using the
“Air Conditioning Load Estimate” form when all
outdoor air is required.
1. Calculate the various loads and determine the
apparatus dewpoint and dehumidified air
quantity.
2. If the dehumidified air quantity is equal to the
outdoor air requirements, the solution is self- .
evident.
Frc.50 - COOL~X ANDDEHUMIDIFICATION
'WITH HIGEI LATENT LOAD
Read t,& where the t&b crosses the straight line plotted
between the outdoor air and room design conditions on
the psychrometric chart, Fig. 50.
t,,b = 6G.6 F
The moisture content at the entering conditions
coil is read from the psychrometric chart.
to
the
w,, = 75.9 gr/lll
7.Determine
leaving conditions of air from cooling coil.
hdb = tadp SBF (kdb - hdp)
(32)
= 48 f.05 (80.4 - 48)
= 49.6
hsa = hadl, + BF (hw - hadp)
(34)
= 19.21 + .05 (31.3 - 19.21)
= 19.82
t[,,,b = 49.1 F
8. Determine supply air temperature to space
RSH
tsa = hn - 1 .os (cfmsa)
(120,000)
= 75 - 1.08 (6940)
=59F
t,,, should also equal
(35)
t,,,,, +
reheat
1.08 (Cf?f&)
W,, =51.1 gr/lb
between room and supply air
=I,.~ - tsa = 75 - 59 = 16
F
=<20 F
9. GTH = 4.45 x 6940 (31.3 - 19.82) = 354,500 Btu/hr
5. Use the recalculated outdoor air loads to determine a new apparatus dewpoint and dehumidified air quantity. This new dehumidified
air quantity should check reasonably close to
the cfmda in 1 tern 1.
A special situation may arise when the condition
explained in Item 4 occurs. This happens when
the ESHF, as plotted on the psychrometric chart.,
does not intersect the saturation line. This situation is handled in a manner similar to that
previously described under “Cooling and Dehumidification -High Latent Load Application.”
Example 3 illustrates an application where codes
specify that al1 outdoor air be supplied to the space.
Example
7 0 2 3 0
= 4g’6 + 1.08 (6940)
= 59 F
Temp. difE
3. If the dehumidified air quantity is less than
the outdoor air requirements, a coil with a
larger bypass factor should be investigated
when the difference in air quantities is small.
If a large difference exists, however, reheat is
required. This situation sometimes occurs
when the application requires large exhaust
air quantities.
.
4. If the dehumidified air quantity is greater than
the outdoor air requirements, substitute cfmda
for cfm,, in the outdoor air load calculations.
(24)
3 - Cooling ond Dehumidification All Outdoor Air
Given:
Application - Laboratory
Location - Wheeling, West Virginia
Summer design - 95 F db, 75 F ~11
Inside design - 75 F db, 55% rh
RSH - 50,000 Btu/hr
RLH - 11,000 Btu/hr
Ventilation - 1600 cfm,,
All outdoor air to be supplied to space.
C H A P T E R 8. AI’I’LIED
1-131
I’SYCHKOMETKICS
times referred to as a “split system.” The moisture
is introduced into the space by using steam or electric humidifiers or auxiliary sprays.
Find:
I.
2.
3.
4.
Outdoor air load (OATH)
Elfective sensil)le heat factor (ESHF)
Apparatus dcwpoint (tad,,)
Dehumidified air quantity (cfm,,J
When humidification is performed in the space,
the room sensible load is decreased by an amount
equal to the latent heat added, since the process is
merely an interchange of heat. The humidifier motor adds sensible heat to the room but the amount
is negligible and is usually ignored.
5. Recalculatetl outtloor air load (OATH)
6. Recalculated effective sensil)lc heat factor (UHF)
7. Final apparatus dewpoint
temperature (tad&
8. Recalculated dehumidified air quantity (cfm&
Solution:
1. OASH = 1.08 X 1600 X (95 - 75) = 34,600 Btu/hr
OALH = 68 X 1600 X (98.5 - 71) = 30,000 Btu/hr
O A T H = 3 4 , 6 0 0 + 30,000 = 64,600 Btu/hr
('4)
(15)
('7)
2. Assume a bypass factor of 0.05 from Tnbles 61 and 62.
ESHF =
=
50,000+(.05)(34,600)
50,000 + (.05) (34,600) + 11,000 + C.05) (30,000)
.81
W)
3. Table 65 shows that, at the given room design conditions
2nd effective sensible heat factor, tadp = 54.5 F.
50,000 +(.05)(34.600)
4. cf?nda= 1.08 (1 - .05) (75 - 54.5) = 2450 cfm
(36)
Since 2450 cfm is larger than the ventilation requirements, and by code all OA is required, the 0.4 loads,
the adp, and the dehumidified air quantity m u s t be
recalculated using 2450 cfm as the OA requirements.
5. Recalculating outdoor air load
OASH = 1.08 X 2450 X (95 - 75) = 53,000 Btu/hr
OALH = .68 X 2450 X (98.5 - 71) = 46,000 Btu/hr
OATH = 53,000 + 46,000 = 99,000 Btu/hr
(14)
(15)
(17)
50,000 + (.05)(53,000)
(50,000) + (.05) (53,000) + 11,000 + (.05) (46,000)
= .80
(26)
6. ESHF=
7. tadp =54F
This checks reasonably close to the value in
.ecalculation is not necessary.
Cooling
With
Step 4, and
Where humidification is required at design to
reduce the air quantity, then a credit to the room
sensible heat should be taken in the amount of the
latent heat from the added moisture. No credit to
the room sensible load is taken when humidification
is usecl to make up a deficiency in the room latent
load during partial load operation.
When the humidifiers and sprays are used to reduce the required air quantity, the latent load
introduced into the space is added to the room
latent load.
When the humidifier or sprays are operated only
to make up the room deficiency, the latent load
introduced into the room by the humidifier or
auxiliary sprays in the space is not added to the ’
room latent load.
The introduction of this moisture into the space
to reduce the required air quantity decreases the
RSHF, ESHF and the apparatus dewpoint. This
method of reducing the required air quantity is
normally advantageous when designing for high
room relative humidities.
The method of determining the amount of moisture necessary to reduce the required air quantity
results in a trial-and-error procedure. The method
is outlined in the following steps:
1.
Humidification
Cooling with humidification may be required at
partial load operation to make up a deficiency in
the room latent load. It may also be used at design
conditions for industrial applications having relatively high sensible loads and high room relative
humidity
requirements. Without
humidification,
excessively high supply air quantities may be required. This not only creates air distribution problems but also is often economically unsound.
Excessive supply air quantity requirements can be
avoided by introducing moisture into the space to
convert sensible heat to latent heat. This is someNOTE: Numbers in parentheses at right edge of column refer
to equations beginning on page 150.
3
3.
Assume an amount of moisture to be added and
determine the latent heat available from this
moisture. Table 64 gives the maximum moisture that may be added to a space without
causing condensation on supply air ducts and
equipment.
Deduct this assumed latent heat from the orignal effective room sensible heat and use the
difference in the following equation for ERSH
to determine tndp.
ndp = t,m -
ERSH
1.08 x (1 - RF) cfmda
is the reduced air quantity permissible
in the air distribution system.
Cfmda
3 . The ESHF is obtained from a psychrometric
chart or Table 65, using the apparatus dewpoint (from Step 2) and room design conditions.
P A R T 1. LOAD ESTIMA+lNG
1-132
4. The new effective room latent load is determined from the Lollowing equation:
EKLH = EKSH x
3. Assume a bypass factor of 0.05 from Tables 61 and 62.
I- ESHF
ESHF
ESHF
The ERSH is from Step 2 and ESHF is from
Step 3.
5. Deduct the original EKLH (before adding
sprays or humidifier in the space) from the new
effective room latent heat in Step 4. The result
is equal to the latent heat from the added
moisture, a n d must check with the value
assumed in Step 1. If it does not check, assume
another value and repeat the procedure.
Example 4 illustrates the procedure for investigating an application where humidification is accomplished within the space to reduce the air
quantity.
lrxample 4 - Cooling With Humidification
Given:
Application - A high humidity chamber
Location - St. Louis, Missouri
Summer design - 95 F db, 78 F WI)
Inside design - 70 F db, 70y0 rh
RSH - 160,000 Btu/hr
RLH - 10,000 Btu/hr
RSHF - .94
Ventilation - 4000 cfrrzoa
in
the
4. Apparatus dewpoint (t&
5. Dehumidified air quantity (cfm,,)
6. Entering and leaving conditions at the apparatus
tcwb’
160.000+(.05)(108.000)
= l?~~0~,~(.05)(108,000)+(.05)(109,000)
= .92
(2~)
4. Plot the ESHF on a psychrometric chart and read the
atlp (dotted line in Fig. 51).
t
a0
= 59.5 F
160,000 + (.05)(108,000) = 15 4oo cfm
’
(36)
= (4000 x 95) i- (1’,400 x 70) = T6,7 F d,)
15,400
(31)
cfrnda = 1.08(1-.05)(70-59.5)
5.
6.
tedb
Read teuzb where the tedb crosses the straight line
plotted between the outdoor and room design conciitions on the psychrometric chart (Fig. 51).
t elcb= 67.9 F wl,
t ldb
= 59.5 + .05 (76.7 - 59.5) = 60.4 F dl,
(32)
Determine the tlwb by drawing a straight line between
the adp and the entering conditions to the apparatus
(the GSHF line). Where tldb intersects this line, read
the tlwb (Fig. 51).
Space
Find:
A. When space humidification is not used:
1. Outdoor air load (OATH)
2. Grand total heat (GTH)
3. Effective sensible heat factor (ESHF)
(tedb’
2. GTH = 160,000 + 10,000 + 108,000 + 109,000
= 387,000 Btu/hr
t Iwb = 60 F WI,
B. When humidification is used in the space:
1. Assume, for the purposeof illustration in this problem,
that the maximum air quantity permitted in the air
distribution system is 10,000 cfm. Assume 5. grains of
moisture per pound of dry air is to be added to convert sensible to latent heat. The latent heat is calcula.ied by m u l t i p l y i n g t h e a i r q u a n t i t y t i m e s t h e
moisture added times the factor .68.
5 X 10,000 X .68 = 34,000 Btu/hr
2. New ERSH = Original ERSH - latent heat of added
moisture
= [lSO,OOO + (.05 X 108,000)] - 34,000
= 131,400 Btu/hr
$db’ bob)
B. When humidification is used in the space:
1. Determine maximum air quantity and assume an
amount of moisture added to the space and latent
heat from this moisture.
2. New effective room sensible heat (ERSH)
3. New apparatus dewpoint (tad&
4. New effective sensible heat factor (ESHF)
5. New effective room latent heat (ERLH)
6. Check calculated latent heat from the moisture added
with amount assumed in Item 1.
7. Theoretical conditions of the air entering the evaporative humidifier before humidification.
8. Entering and leaving conditions at the apparatus
edb’ tezabJ t,db’ $,ab)
Solution:
A. When space humidification is not used:
1. OASH = 1.08 X 4000 X ( 95-50) = 108,000 Btu/hr (14)
OALH = 68 x 4000 x (117-77) = 109,000 Btu/hr (15)
= 217,000 Btu/hr (17)
OATH= 108,000 + 109,000
NOTE: Numbers in parentheses at right edge of column refer
to equations beginning on page 150.
3.
tadp
131,400
= ” - 1.08 (1 - .05) (10,000) = 57’2 F
(36)
4 . ESHF is read from the psychrometric chart as .73
(dotted line in Fig. 52).
>
I- ESHF
5. New ERLH = New ERSH X
ESHF
1 - .73
= 131,400 x ~
.73
= 48,600 Btu/hr
6. Check for latent heat of added moisture.
Latent heat of added moisture
= New ERLH - Original ERLH
= 48,600 - [IO,000 + (.05 X 109,000)]
= 33,200 Btu/hr
This checks reasonably close
in Step 2 (34,000 Btu/hr).
i. Psychrometrically,
it can
ized water from the spray
part of the room sensible
vapor at the final room
theoretical dry-bulb of the
with
the
assumed
value
be assumed that the atomheads in the space absorbs
heat and turns into water
wet-bull, temperature. The
air entering the sprays is
,
at the intersection of the
anti the moisture content of
‘I‘his moisture content is
the moisture atltletl I)y the
design moisture content.
room design wet-Ilull) line
the air entering the sprays.
tlcrerminetl I)y sril)tracting
room sprays Irom the room
Moisture content of air entering
= 77 - 5 = 52 p/lb.
humitlilier
‘I’hc
theorctic;~l tlry-bull) i s tlctcrminctl from t h e
pychromctric chart x i3.3 (II), illustratctl on I;ig. 52.
8. t,,d,, =
(4000 x 9.5) + (6000 x i0) = 8. F (,,,
10,000
(31)
where the Ipdh crosses the straight line
Read l<,,,.b
plotted between the outdoor and room design con.
tlitions on the psychrometric chart (Fig. 52).
QMFdb
70Fdb
76.7 F db
95 Fdb
FIG. 51 -COOLING
AND
DEHUMIDIFICATION
AI)DING No MOISTURETOTHE SPACE
t
F”, b
t ldh
= 69.8
I< WI)
= 57.2 + (.05)(80 - 57.2) = 58.4 1: tll,
(8’2)
Determine tlu,* by drawing a straight line Ixtween
the atlp and the entering conditions to the apparatus
(GSHF line). Where tld,, intersects this line, read the
llw,, (Fig. 52).
t
lwb = 58 F wh
The straight line connecting the leaving conditions at the apparatus with the theoretical condition
of the air entering the evaporative humidifier rcpresents the theoretical process line of the air. This
theoretical condition of the air entering the humidifier represents what the room conditions are it’ the
humidifier is not operating. The slope of this theorctical process line is the same as RSHF (.94).
The heavy lines on Fig. 52 illustrate the theoretical air cycle as air passes through the conditioning apparatus to the evaporative humidifier, then
to the room, and finally back to the apparatus where
the return air is mixed with the ventilation air.
Actually, if a straight line were drawn from the
leaving conditions of the apparatus (58.4 F db, 58
F wb) to the room design conditions, this line would
be the RSHF line and would be the process line for
the supply air as it picks up the sensible and latent
loads in the space (including the latent heat added
by the sprays).
The following two methods of. laying out the
system are recommended when the humidifier is to
be used for both partial load control and reducing
the air quantity.
1. Use two humidifiers; one to operate continuously, adding the moisture to reduce the air
quantity, and the other to operate intermittently to control the humidity. The humidifier
used for partial load is sized for the effective
room latent load, not including that produced
by the other humidifier. If the winter requirements for moisture addition are larger than
summer requirements, then the humidifier is
selected for these conditions. This method of
using two humidifiers gives the best control.
F IG.
52
ADDING
-COOLING
MOISTURE
DEHUMIDIFICATION
INTO THE SPACE
AND
2. Use one humidifier of sufficient capacity to
handle the effective room latent heat plus the
calculated amount of latent heat from the
added moisture required to reduce the air
quantity. In Part B, Step 5, the humidifier
would be sized for a latent load of 48,600
Btu/hr.
I’AK?‘
1-134
Sensible Cooling
A sensible cooling process is one that removes
heat from the air at a constant moisture content,
line (I - 2), Fig. 48. Sensible cooling occurs when
either of the following conditions exist:
1. LOAD ESTIMATING
Example 5 illustrates the method of determining
the effective surface temperature for a sensible cooling application.
Example 5 - Sensible Cooling
Given:
1. The ESHF calculated on the air conditioning
load estimate form is equal to 1.0.
OT
2. The entering and leaving conditions at the
apparatus, as checked or plotted on the psy
chrometric chart, indicate a GSHF equal to 1 .O.
In a sensible cooling application the ESHF, GSHF
and RSHF all equal 1.0. When the RSHF equals
1.0, however, it does not necessarily indicate a sensible cooling process because latent load, introduced
I outdoor air, can give a GSHF less than 1.
The apparatus dewpoint is referred to as the
effective surface temperature (tes) in sensible cooling
applications. The effective surface temperature
must be equal to, or higher than, the dewpoint
temperature. In most instances, the t,, does not lie
on the saturation line and, therefore, will not be
the dewpoint of the apparatus. Whether or not t,,
lies on the saturation line depends entirely on the
bypass factor of the coil selected for the application.
However, the calculations for ESHF, adp and cfmda
may still be performed on the air conditioning load
estimate form by substituting the term t,, for tadp.
The use of the term cfmda in a sensible cooling application should not be construed to indicate that
dehumidification is occurring. It is used in the “Air
Conditioning Load Estimate” form and in Example
5 to determine the air quantity required thru the
apparatus to offset the conditioning loads.
~, The leaving air conditions from the coil are dictated by the room design conditions, the load and
the required air quantity. The entering and leaving
dry-bulb temperatures at the apparatus should always be used to determine the effective surface
temperature of the coil. Since this is strictly a
sensible heat process, it is straight line function
occurring at constant moisture content. Introducing
wet-bulb into the calculation results in an erroneously low t,,. If this erroneous t,, is used to select
the conditioning equipment, then 1. Direct expansion equipment would be selected
at a lower refrigerant temperature than actually required.
2. Chilled water equipment would be selected
for colder or more chilled water than actually
required.
.
Location - Bakersfield, California
Summer design - IO5 F dh, 70 F WI)
Inside design - 75 F dh, 50’;:, maximum rh
RSH - 200,000 Btu/hr
RLH - none
Ventilation - 13,000 c/m,,,
Find:
1.
2.
3.
4.
5.
6.
Outdoor air load (OATH)
Effective sensible heat factor (ESHF)
Effective surface temperature (tea)*
Dehumidified air quantity (cfm,,)
Effective surface temperature (tes)*
Supply air temperature (tJ
Solution:
OASH = 1 . 0 8 X 1 3 , 0 0 0 (105 - 75) = 420,000 Btu/hr
(14)
OALH = .68 X 1 3 , 0 0 0 (54 - 54) = 0
OATH = 420.000 Btu/hr
(15)
(17)
Assume a bypass factor of 0.05 from Tables 61 and 62.
200,000 + (.05) (420,000)
ESHF = 200,000 + (.05) (420,000) = “ ’
Plot the ESHF to the saturation line on the psychrometric chart. The effective surface temperature is read
as tcs = 50.3 F, Fig. 53.
200,000 + (.05) (420,000)
cf “Ida = 1.08 x (1 - .05) X (75 - 50.3) = 8650 cfm
36)
Since the dehumidified air quantity is less than the outdoor ventilation requirements, substitute the cfn,,
for
in
Step
4.
This
results
in
a
new
effective
surface
Cfnda
temperature which does not lie on the saturation line.
This is illustrated in the following equation:
tes
= 75 -
200,000 + (.05) (420,000)
1.08 X (1 - .05) X 13,000
= 58.4 F
(36)
This temperature falls on the ESHF line, which is also
the GSHF and RSHF lines in a sensible cooling appli4
cation.
6. Substitute tea for tadp in equation (28),
supply air condition tsa as follows:
t 8LI = 105 - (1 - .05) (105 - 58.4)
tda is the same as the tldb which
and calculate the
= 60.7 F db
(28)
is the leaving air condition from the coil. The wet-bulb temperature of the
air supplied to the space is 54 F wb. This is read at the
intersection of the supply air dry-bull, temperatnre and
the ESHF line, Fig. 53.
.
In Example 5, the assumed .05 bypass factor is
used to determine t,, and dehumidified air quantity.
Since the dehumidified air quantity is less than
*The terms tep is substituted for tadp on the air conditioning
load estimate‘form for sensible cooling application.
NOTE: Numbers in parentheses at right edge of column refer
to equations heginning on page 150.
CHAI’I‘EK
8. AI’I’LIEI~
1-135
I’SYCI-IROME’I‘RICS
ventilation air rcquircment, the .05 bypass factor
is used again to determine a new t,,, substituting
the ventilation air requirement for the dchumidified air quantity. The new L,, is 58.4 F.
ciency. It can be considered to represent that portion
of the air passing thru the spray chamber which
contacts the spray water surface. This contacted air
is considered to be leaving the spray chamber at
the effective surface temperature of the spray water.
This effective surface temperature is the temperature at complete saturation of the air.
Though not a straight line function, the e,rcct
of saturation efficiency on the leaving air conditions
from a spray chamber may be determined with a
sufficient degree of accuracy from the following
equation:
54 gr/lb
Ll
58.4 F db 64.7 Fdb
75Fdb
105 Fdb
FIG . 53 - S ENSIBLE COOLING
If a coil with a higher bypass factor is substituted in Example 5, a lotier t,, results. Under
these conditions, it becomes a question of economic
balance when determining which coil selection and
which refrigerant temperature is the best for the
application. For instance, the maximum possible
coil bypass factor that can be used is .19. This still
results in a t,, above 50.3, and at the same time
maintains a dehumidified air cfm of 13,000 which
equals the ventilation requirements.
SPRAY CHARACTERISTICS
J, the operation of spray type equipment, air
is c-,.wn or forced thru a chamber where water is
sprayed thru nozzles into the air stream. The spray
nozzles may be arranged within the chamber to
spray the water counter to air flow, parallel to air
flow, or in a pattern that is a combination of these
two. Generally, the counter-flow sprays are the most
efficient; parallel flow sprays are the least efficient;
and when both are employed, the efficiency falls
somewhere in between these extremes.
SATURATION
EFFICIENCY
In a spray chamber, air is brought into contact
with a dense spray of water. The air approaches the
state of complete saturation. The degree of saturation is termed saturation efficiency (sometimes called
contact or performance factor). Saturation efficiency
is, therefore, a measure of the spray chamber effi-
The saturation efficiency is the complement of
bypass factor, and with spray equipment the bypass
factor is used in the calculation of the cooling load.
Bypass factor, therefore, represents that portion of
the air passing thru the spray equipment which is
considered to be leaving the spray chamber completely unaltered from its entering condition.
This efficiency of the sprays in the spray chamber
is dependent on the spray surface available and on
the time available for the air to contact the spray
water surface. The available surface is determined
by the water particle size in the spray mist (pressure
at the spray nozzle and the nozzle size), the quantity
of water sprayed, number of banks of nozzles, and
the number of.nozzles in each bank. The time available for contact depends on the velocity of the air
thru the chamber, the length of the effective spray
chamber, and the direction of the sprays relative to
the air flow. As the available surface decreases or as
the time available for contact decreases, the saturation efficiency of the spray chamber decreases. Table
63 illustrates the relative efficiency of different spray
chamber arrangements.
The relationship of the spray water temperatures
to the air temperatures is essential in understanding
the psychrometrics of the various spray processes. It
can be assumed that the leaving water temperature
from a spray chamber, after it has contacted the air,
is equal to the leaving air wet-bulb temperature.
The leaving water temperature will not usually vary
more than a degree from the leaving air wet-bulb
temperature. Then the entering water temperature
is, therefore, dependent on the water quantity and
the heat required to be added or removed from
the air.
Table 63 illustrates the relative efficiency of different spray chamber arrangements.
1-136
TABLE
I’:\K’l‘
63-TYPICAL SATURATION
I. IDi\D
ESTIMi\‘l‘lNG
EFFICIENCY*
For Spray Chambers
l/q”
NO.
OF
BANKS
DIRECTION
OF
WATER
SPRAY
NOZZLE
(25 psig
Nozzle Pressure
3 gem/sq
ftt)
Velocityf
I/~”
NOZZLE
(30 psig
Nozzle Pressure
2.5 gpm/sq ftt)
(fpm)
300
700
300
700
1
Parallel
Counter
70%
75%
50%
65%
80%
82%
60%
70%
2
Parallel
Opposing
Counter
90%
98%
99%
85%
92%
93%
92%
98%
99%
87%
93%
94%
‘Saturation efficiency = 1 - BF
tCpm/sq
ft of chamber face area
TVelocities above 700 fpm and below 300 fpm normally do
not permit eliminators to adequately remove moisture from
.ie air. Reference to manufacturers’ data is suggested for
.miting velocity and performance.
SPRAY PROCESSES
Sprays are capable of cooling and dehumidifying,
sensible cooling, cooling and humidifying, and
heating and humidifying. Sensible cooling may be
accomplished only when the entering air dewpoint
is the same as the effective surface temperature of
the spray water.
The various spray processes are represented on
the psychrometric chart in Fig. 54. All process lines
must go toward the saturation line, in order to be
at or near saturation.
Adiabatic Saturation or Evaporative Cooling
Line (1 - 2) represents the evaporative cooling
process. This process occurs when air passes thru
a spray chamber where heat has not been added to
-.r removed from the spray water. (This does not
.clude heat gain from the water pump and thru
the apparatus casing.) When plotted on the psychrometric chart, this line approximately follows
up the line of the wet-bulb temperature of the air
entering the spray chamber. The spray water temperature remains essentially constant at this wetbulb temperature.
Cooling and Humidification - With Chilled Spray Water
If the spray water receives limited cooling before
it is sprayed into the air stream, the sIope of the
process line will move down from the evaporative
cooling line. This process is represented by line
(I - 3). Limited cooling causes the leaving air to be
lower in dry- and wet-bulb temperatures, but higher
in moisture content, than the air entering the spray
chamber.
.
.
DRY-BULB TEMPERATURE
FIG .
54 - S PRAY P ROCESSES
Sensible Cooling
If the spray water is cooled further, sensible
cooling occurs. This process is represented by line
(1 - 4). Sensible cooling occurs only when tho entering air dewpoint is equal to the effective surface
temperature of the spray water; this condition is
rare. In a sensible cooling process, the air leaving
the spray chamber is lower in dry- and wet-bulb
temperatures but equal in moisture content to the
entering air.
Cooling and Dehumidification
If the spray water is cooled still further, cooling
and dehumidification takes place. This is illustrated
by line (1 - 5). The leaving air is lower in dry- and
wet-bulb temperatures and in moisture content
than the air entering the spray chamber.
Cooling and Humidification - With Heated Spray Water
When the spray water is heated to a limited degree before it is sprayed into the air stream, the
slope of the process line rises to a point above the
evaporative cooling line. This is illustrated by line
(1 - 6). Note that the leaving air is lower in dry-bulb
temperature, but higher in wet-bulb temperature
and moisture content, than the air entering the spray
chamber.
Heating and Humidification
If the spray water is sufficiently heated, a heating
and humidification process results. This is represented by line (1 - 7). In this process the dry-bulb
CH,\I”I‘EK 8 . AI’I’LIEI)
temperature, wet-bulb temperature, and moisture
content of the leaving air is greater than that of the
entering air.
SPRAY PROCESS EXAMPLES
The following descriptions and examples provide
a better understanding of the various psychrometric
processes involved in spray washer equipment.
Cooling and Dehumidification
w h e n
a
1-137
I’SYCHKOME’I’KICS
S p r a y chanlberiS t o
b e used h coohg
and dehumidification, the procedure for estimating
the load and selecting the equipment is identical
to the procedure described on page 128 for coils.
The “Air Conditioning Load Estimate” form is used
to evaluate the load; bypass factor is determined by
subtracting the selected saturation efficiency from
one. Spray chamber dehumidifiers may not be rated
in ’ -ms o f a p p a r a t u s dewpoint but in terms of
ent. -“g arid leaving wet-bulb temperatures at the
apparatus. The apparatus dewpoint must still be
determined, however, to evaluate properly the entering and leaving wet-bulb temperatures and the
dehumidified air quantity.
Although originally prepared to exemplify the
is also
operation of a coil,
typical of the cooling and &humidifying process
using sprays.
Cooling and Dehumidification - Usiqg
All Outdoor Air
When a spray chamber is to be used for cooling
and dehumidifying with all. outdoor air, the procedure for determining adp, entering and leaving
conditions at the chamber, ESHF and cfmda is
identical to the procedure for determining these
items for coils using all outdoor air. Therefore, the
description on page 130 and Example 3 may be used
to ’
lyze this type of application.
temI)erature is to he maintained during the winter
or intermediate season, heat must be available to
the system. This is usually accomplished by adding
a reheat coil. When relative humidity is to be maintained in addition to room dry-bulb during the
winter or intermediate season, a combination of
preheat and reheat coils, or a reheat coil and spray
water heating, is required. The latter method
changes the process from evaporative cooling to
one of the humidification processes illustrated by
lines (1 - 6) or (I - 7) in Fig. 54.
Evaporative cooling may be used in industrial
applications where the humidity alone is critical,
and also in dry climates where evaporative cooling
gives some measure of relief by removing sensible
heat.
L;xnmple 6 illustrates an industrial application
designed to maintain the space relative humidity
only.
Example 6 - Evaporative Cooling
Given:
An industrial application
Location - Columbia, South Carolina
Summer design - 95 F db. 75 F wh
Inside design - 55~~ rh
RSH - ?,lOO.OOO Btu/hr
RSHF - 1.0
Use all outdoor air at design load conditions
Find:
1. Room dry-bulb temperature at design (f,,)
2. Supply air quantity (cftn,,)
Evaporative Cooling
An evaporative cooling application is the simultaneous removal of sensible heat and the addition
of moisture to the air, line (1 - 2), Fig. 54. The spray
water temperature remains essentially constant at
the wet-bulb temperature of the air. This is a process
in which heat is not added to or removed from the
spray water. (Heat gain from the water pump and
heat gain thru the apparatus casing are not included.)
Evaporative cooling is commonly used for those
applications where the relative humidity is to be
controlled but where no control is required for the
room dry-bulb temperature, except to hold it above
a predetermined minimum. When the dry-bulb
b-92.3 Fdb C-94.0 F db
B-93.2 F db D-95.0 F db
ABC0
FIG. 55 - EVAPORATIVE COOLING, W ITH VARYING
SATURATION EFFICIENCY
1-138
PAKT
I.
LO:\D E S T I M A T I N G
L
Solution:
I . Dcterminc the room dry-l)ulb temperature by comp r o m i s i n g I)ctwcen the spray saturation efficiency, the
acccl~tablc room dry-l)ull) tempcraturr, and the supply
air quantity. To evaluate these items, use the following
equation to tlctcrrninc
the leaving conditions from the
spray for variorls saturation efficiencies:
1
te d b - (Sat Elf) (lEdb - tczub)”
Tllc I-oom dry-l~ulb tcrnperatllrc
in the following table
results from various spray saturation efficiencies and is
dctennined
by plotting the RSHF thru the various leaving conditions, to the design relative humidity, Fig. 55.
Note that the supply air temperature rise clecreases more
rapidly than the room dry-bulb temperature. Correspondingly,
as the supply air temperature rise decreases, the supply air quantity increases in t h e s a m e
proportion.
DRY-BULB
TEMP
LEAVING
SPRAYS
S.4T
EFF
(%o)
(tldh)
2.
100
75
95
90
85
80
76
77
78
79
SUPPLY
AIR
TEMP
RISE
(At)
with the auxiliary sprays in the space, becomes a
problem of economics which shouitl be analyzed
for each particular application.
When a split system is used, supplemental spray
heads arc usually added to the straight evaporative
cooling system. These spray heads atomize water
and add supplementary moisture directly to the
room. This added moisture is evaporated at the
final room wet-bulb temperature, and the room
sensible heat is reduced by the amount of heat
required to evaporate the sprayed water.
Table 64 gives the recommended maximum moisture to be added, based on a 65 F db room temperature or over, without causing condensation on the
ductwork.
ROOM
DRY-BULB
TEMP
AT 55% RH
it,,)
I9
17.6
16.2
14.7
13.3
various
Without
ROOM
DESIGN
RH
94
93.6
93.2
92.7
92.3
Calculate the supply air quantity for the
perature rises from the following equation:
TABLE 64-MAXIMUM RECOMMENDED
MOISTURE ADDED TO SUPPLY AIR
85
80
75
tem-
Causing
/MOISTURE
Gr/Cu Ft
Dry Air
1.25
1.30
1.35
1.40
70
Condensation
on
Ductct
1ROOM / MOISTURE
DESIGN
Gr/Cu Ft
RH
Dry Air
65
60
55
1.50
1.60
1.70
1.80
50
tThese are arbitrary limits which have been established by a
combination of theory and field experience. These limits apply where the room dry-bulb temperature is 65 F db or over.
SUPPLY AIR
TEMP RISE
(tWTl
SUPPLY AIR
QUANTITY
- ‘Mb)
(cfm,,)
19
102,400
The spray chamber and supply air quantity should
then be selected to result in the best owning and operating costs. The selection is based primarily on economic
considerations.
Evaporative Cooling Used With A Split System
There are occasions when using straight evaporative cooling results in excessive air quantity requirements and an unsatisfactory air distribution system.
This situation usually arises in applications that
are to be maintained at higher relative humidities
(70% or more). To use straight evaporat’ve cooling
with the large air quantity, or to use a s t lit system
*This equation is applicable only to evaporative cooling applications where the entering air wet-bulb temperature, the
leaving air wet-bulb temperature, and the entering and leaving water temperature to the sprays are all equal.
.
4s a rule of thumb, the air is reduced in temperature approximately 8.3 F for every grain of moisture
per cubic foot added. This value is often used as
a check on the final room temperature as read
from the psychrometric chart.
Example 7 illustrates an evaporative cooling application with supplemental spray heads used in
the space.
Example
74voporafive
Cooling-With
Auxiliary
Sprays
Given:
An industrial application
Location - Columbia, South Carolina
Summer design - 95 F db, 75 F wb
Inside design - TOY0 rh
RSH - 2.100.000 Btu/hr
RSHF - 1.0
Moisture added by auxiliary sp’ray
heads - 19 gr/lb (13.9
cu ft/lb X 1.4 gr/cu ft)
Use all outtloor air thru a spray chamber with 90y0 saturation efficiency.
Find:
1.
2.
3.
4.
Leaving
Room
Supply
Supply
conditions from spray chamber pldb, tIwb)
dry-bulb temperature (tr,,)
air quantity (cfm,,) with auxiliary sprays
air quantity (cfm,,) without auxiliary sprays
,
1-139
CH,\I'J‘EK 8. i\l'I'L.Il~I~ I'SYCHKOME'I‘KICS
Cfffl,, = -- 1.08 X
F
EVAP COOLING
RSH
t-i&
temp
=
2,100,000
1.08 X 23.75 - = 82,000 cfm
4. If no auxiliary sprays were to I)e used, the room design
dry-l)ulb
would be where the RSHF line intersects the
room tlesign relative humidity. From Fig. 56, the room
dry-bulb is read
t rn”A - 84.7 I; tlb
The supply air quantity required to maintain the room
design relative humidity is determined from the following
equation:
\_THE~RETICAL
db TEMP
RISE
WITHOUT
AUXILIARY
SPRAYS
cf ItIle =
Heating
FIG.
56
89.2 Fdb 95Fdb\
94.7F db
100.75F db
-EVAPORATIVE COOLING, WITH AUXILIARY
SPRAYS WITHINTHE
SPACE
Solution:
1. tldb = ted,-- (Sat Eff) Ctedb - tewb)
= 95 - .90 (95 - 75) = 77 F db
t Iwb is the same as the tewb
process, Fig. 56.
2.
in an evaporative cooling
Room dry-bulb temperature is evaluated
the moisture content of the space.
W rm = Wsa
by
determining
+ 19 = 128 + 19 = 147 gr/lb
The 19 gr/lb is the moisture added to the space by the
auxiliary spray heads.
T h e trm is the point on the psychrometric chart where
h e
Wrm intersects the 70% design relative humidity
line, Fig. 56.
t
t
3.
rm
= 89.2 F db
Psychrometrically,
it can be assumed that the atomized
water from the spray heads absorbs part of the room
sensible heat and turns into water vapor at the final
room wet-bulb temperature. The intersection of this wetbull, temperature with the moisture content of the air
leaving the evaporative cooler is the theoretical dry-bulb
equivalent temperature if the auxiliary sprays were not
operating. The difference between this theoretical drybulb equivalent temperature and the temperature of the
spray chamber, tldb, is used to determine the supply air
quantity.
tldO
(from spray chamber) = 77 F.
The theoretical dry-bulb temp is 100.75
Temp rise = 23.75 F dh
2.100,000
= 1.08 (84.7 - 77)
= 253,000 cfm
This air
required
However,
quantity,
84.7 F to
I
77Fdb
RSH
1.08 Cfrm - tld*l
I;, Fig. 56.
and
quantity is over three times the air quantity
when auxiliary sprays are used in the space.
it should be noted that, by reducing the air
the room dry-bulb temperature increased from
89.2 F.
Humidification
-With
Sprays
A heating and humidifying application is one in
which heat and moisture are simultaneously added
to the air, line (1 - 7), Fig. 54. This may be required
during the intermediate and winter seasons or during partial loads where both the dry-bulb tempera- /
ture and relative humidity are to be maintained.
Heating and humidification may be accomplished
by either of the following methods:
1. Add heat to the spray water before it is
sprayed into the air stream.
2. Preheat the air with a steam or hot water
coil and then evaporatively cool it in the spray
chamber.
Spray water is heated, by a steam to water interchanger or by direct injection of steam into the
water system. Since the supply air quantity and the
spray water quantity have been determined from
the summer design conditions, the only other requirement is to determine the amount of heat to
be added to the spray water or to the preheater.
For applications requiring humidification, the
room latent load is usually not calculated and the
room sensible heat factor is assumed to be 1 .O.
Example 8 illustrates the psychrometric calculations for a heating and humidifying application
when the spray water is heated. It should be noted
that this type of application occurs only when the
quantity of outdoor air required is large in relation
to the total air quantity.
Example 8 - Heating and Humidification With Heated Spray Water
Given:
4n industrial application
Location - Richmond, Virginia
l-140
PART
the wet-hulh
line, Fig. 54.
Winter tlesign - 15 F dh
Inside design - 72 F tll), 35”/” rh
Ventilation - 50,000 cffnon (see explanation above)
Supply air - 85,000 ~/rn,~
Design room heat loss - 2,500,OOO Btu/hr
Spray saturation efficiency - 95%
RSHF (winter conditions) - 1.0
Make-up water - 65 F
t
1. tsn
=
~~xnt =
wkI - wea
Sat Eff
+ We=a
41 - 17
=- + 17 = 42.3 gr/lb
.95
The heating and humidification process line is plotted
on the psychrometric chart between the moisture content
of saturated air (42.3 gr/lb) and the entering conditions
to the spray chamber (38.5 F db and 32.4 F wb), Fig. 57.
+ t,n,
To determine the wet-bulb temperature, plot the RSHF
line on the psychrometric chart and read the wet-bulb
at the point where t,@ crosses this line (Fig. 57). Supply
air wet-bull, to the space = 65.8 F wb.
The leaving conditions
chart where the room
intersects the heating
Fig. 57.
= 38,5 F d,,
(15 X 50,000) + (72 x 35,000)
85,000
-.-
tlwb = 43.4 F wb
The temperature of the leaving spray water is approximately equal to the wet-bulb temperature of the air
leaving the spray chamber.
t
Iw = 43.4 F
(31)
To determine wet-bull) temperature of the air entering
the spray chamber, plot the mixture line of outdoor and
return room air on the psychrometric chart, and read
are read from the psychrometric
moisture content line (41 gr/lb)
and humidification process line,
t Idb = 43.6 F db
2. To determine the entering and leaving spray water
temperature, calculate the entering and leaving air conditions at the spray chamber:
=
= 32.4 F WI,
tllL.)
2,500,000
+72=99.2 F d h
= 1.08 X 85,000
t edb
NOTE: Numbers in parentheses at right edge of column. refer
to equations heginning on page IJO.
SUPPLY
w.. .-
AIR
I/
REHEAT
42.3 y/lb
41 grllb
1 7 gr/lb ’
I3 Fdb
FIG. 57
crosses the mixture
W )T,L = Wla = 41 gr/lh
Since the spray chamber
has a saturation efficiency of
‘3~5%~ the moisture content of completely saturated air
is calculated as follows:
Solution:
design room heat loss
1.08 x ~ffn,~
temperature where tedb
The air leaving the spray chamber
must have the same
moisture content as the air in the room.
Find:
1. Supply air conditions to the space (t3
2. Entering and leaving spray water temperature jtCw,
3. Heat added to spray water to select water heater,
rwb
I. L O A D E S T I M A T I N G
38.5 Fdb 4L6 Fdb
-
HEATING
AND
.
?ZFdb
99.2 Fdb
HUMIDIFICATION, W ITH HEATING S PRAY
W ATER
l
1-141
CHAPTER 8. APPLIED I’SYCHROMETRICS
The
temperature of the entering spray water is dependent
on the water quantity and the heat to Ix! added or
removed from the air. In this type of application, the
water quantity is us~~ally dictated hy the cooling t&l
design requirements. ,\ssume,
for illustration purposes,
that this spray washer is selected for 110 gpm for
cooling.
The heat added to the air as it passes through the washer
= cfn,, x 4.45 x (llL1. - ‘2,J
= 85,000 X 4.45 X (16.85 - 12)
= 1,830,OOO Btu/hr
The entering water
following
equation:
t ew
= tlw
temperature
is
determined
from
the
heat added to air
500 x gpm
+
I ,830,000
= 43.4 + 500 x 110
= 76.8 F
3. The heat added to the spray water (for selecting spray
-%er heater) is equal to the heat added to the air plus
heat ,added to the make-up water. The amount of
make-up water is equal to the amount of moisture evaporated into the air and is determined from the following
equation:
outdoor air and mixing it with the return air from
the space. This mixture must then be evaporativcly
cooled to the room dewpoint (or room moisture
content). And finally, the air leaving the spray
chamber must be reheated to the required supply
air temperature.
SORBENT DEHUMIDIFIERS
Sorbent dehumidifiers contain liquid absorbent
or solid adsorbent which are either sprayed directly into, or located in, the path of the air stream.
The liquid absorbent changes either physically or
chemically, or both, during the sorption process.
The solid adsorbent does not change during the
sorption process.
As moist air comes in contact with either the
liquid absorbent or solid adsorbent, moisture is
removed from the air by the difference in vapor
pressure between the air stream and the sorbent. As
cfm,, P1, - WcJ
Make-up
where:
water =
wea, Wza
7000
12.7
8.34
7000 X 12.7 X 8.34
I
= moisture content of the air entering and
leaving the spray washer in grains per
pound of dry air
= grains of moisture per pound of dry air
= volume of the mixture in cubic feet per
pound of dry air, determined from psychrometric chart
= water in pounds per gallon
85,000 (41 - 17)
Make-up water = 7ooo x 12.7 x 8.34 = 2.8 gpm
The heat added to the make-up spray water is determined
from the following equation:
Heat added to make-up water
.
= gpm X 500 (tew - make-up water temp)
= 2.8 X 500 (76.8 - 65)
= 16,200 Btu/hr
To select a water heater, the total amount of heat added
to the spray water is determined by totaling the heat
added to the air and the heat added to the make-up
spray water.
Heat added to spray water
= 1,830,OOO + 16200
= 1,846,200 Btu/hr
If the make-up water was at a higher temperature than
the required entering water temperature to the sprays,
then a credit to the heat added to the spray water may
be taken.
In this example a reheat coil is required to heat
the air leaving the spray chamber, at 43.6 F db and
at a constant moisture content of 41 gr/lb, to the
required supply air temperature of 99.2 F db.
The requirements of the application illustrated
in Example 8 can also be met by preheating the
DRY-BULB
F IG.
TEMPERATURE
58 - SORBENT DEHUMIDIFICATION PROCESSES
this moisture condenses, latent heat of condensation is liberated, causing a rise in the temperature
of the air stream and the sorbent material. This
process occurs at a wet-bulb temperature that is
approximately constant. However, instead of adding moisture to the air as in an evaporative cooling
process, the reverse occurs. Heat is added to the air
and moisture is removed from the air stream; thus
it is a dehumidification and heating process as illustrated in Fig. 58. Line (I - 2) is the theoretical
process and the dotted line (I -3) approximates what
actually happens. Line (I - 3) can vary, depending
on the type of sorbent used.
I’AK-I‘
1-142
I. LOAD ESI’IM,\~I‘ING
PSYCHROMETRICS OF PARTIAL LOAD CONTROL
The apparatus required to maintain proper
space conditions is normally selected for peak load
operation. Actually, peak load occurs but a few
times each year and operation is predominantly
at partial load conditions. Partial load may be
caused by a reduction in sensible or latent loads in
the space, or in the outdoor air load. It may also be
caused by a reduction in these loads in any combination.
PARTIAL LOAD ANALYSIS
Since the system operates at partial load most
of the time and must maintain conditions commensurate with job requirements, partial load
analysis is at least as important as the selection
of equipment. Partial load analysis should include
study of resultant room conditions at minimum
total load. Usually this will be sufficient. Certain
applications, however, should be evaluated at minimum latent load with design sensible load, or
minimum sensible load and full latent load. Realistic minimum and maximum loads should be
assumed for the particular application so that, psychrometrically, the resulting room conditions are
properly analyzed.
Figure 57 illustrates the psychrometrics of reheat
control. The solid lines represent the process at
design load, and the broken lines indicate the
resulting process at partial load. The RSHF value,
plotted from room design conditions to point (2),
must be calculated for the minimum practical room
sensible load. The room thermostat then controls the temperature of the air leaving the reheat
coil along line (1 - 2). This type of control is applicable for any RSHF ratio that intersects line (I - 2).
If the internal latent loads decrease, the resulting
room conditions are at point (3), and the new RSHF
process line is along line (2 - 3). However, if humidity is to be maintained within the space, the
reduced latent load is compensated by humidifying,
thus returning to the design room conditions.
The six most common methods, used singly or in
combination, of controlling space conditions for
cooling applications at partial load are the following:
(SENSIBLE
1. Reheat the supply air.
2. Bypass the heat transfer equipment.
3. Control the volume of the supply air.
4. Use on-off control of the air handling equipment.
HEATING)
\
‘NEW RSHF LINE i
WITH REDUCED ROOM
SIGN
DRY-BULB
I
SENSIBLE HEAT
J
DRY-BULB
TEMPERATURE
5. Use on-off control of the refrigeration machine.
6. Control the refrigeration capacity.
The type of control selected for a specific application depends on the nature of the loads, the conditions to be maintained within the space, and
available plant facilities.
REHEAT CONTROL
Reheat control maintains the dry-bulb temperature within the space by replacing any decrease in
the sensible loads by an artificial load. As the internal latent load and/or the outdoor latent load
decreases, the space relative humidity decreases. If
humidity is to be maintained, rehumidifying is reqy.ired
in addition to reheat. This was described
previously under “Spray Process, Heating and
Humidifying.”
.
FIG. 59 - PSYCHROMETRICS
OF
REHEAT CONTROL
BYPASS CONTROL
Bypass control maintains the dry-bulb temperature within the space by modulating the amount of
air to be cooled, thus varying the supply air temperature to the space. Fig. 60 illustrates one method
of bypass control when bypassing return air only.
Bypass control may also be accomplished by
bypassing a mixture of outdoor and return air
around the heat transfer equipment. This method
of control is inferior to bypassing return air only
since it introduces raw unconditioned air into the
space, thus allowing an increase in room relative
humidity.
.
1-143
CHAPTICK 8. AI'I'LIEI) I'SY(:l-IKO~lI1'TKICS
DESIGN
SUPPLY CON&
DESIGN
DESIGN RSHF
NEW RSHF LINE WITH
REDUCED ROOM
SENSIBLE HEAT.
AROUNO APPARATUS
MIXTURE OF BYPASSED AIR
AND AIR THRU OEHUMIDIFIER
DRY-BULB TEMPERATURE
FIG. 60 -PSYCHROMETRICSOFBYPASSCONTROLWITH
A reduction in room sensible load causes the
bypass control to reduce the amount of air thru the
dehumidifier. This reduced air quantity results in
equipment operation at a lower apparatus dewpoint. Also, the air leaves the dehumidifier at a
1C
- temperature so that there is a tendency to
ac,.“st for a decrease in sensible load that is proportionately greater than the decrease in latent
load.
Bypass control maintains the room dry-bulb temperature but does not prevent the relative humidity
from rising above design. With bypass control,
therefore, increased relative humidity occurs under
conditions of decreasing room sensib!e
load and
relatively constant room latent load and outdoor
air load.
The heavy lines in Fig. 60 represent the cycle for
design conditions. The light lines illustrate the
initial cycle of the air when bypass control first
begins to function. The new room conditions, mixture conditions and apparatus dewpoint continue
KETURNAIRONLY
to change until the equilibrium point is reached.
Point (2) on Figs. 60 and 61 is the condition of
air leaving the dehumidifier. This is a result of
a smaller bypass factor and lower apparatus dewpoint caused by less air thru the cooling equipment
and a smaller load on the equipment. Line (2 - 3 - 4)
represents the new RSHF line caused by the reduced
room sensible load. Point (3) falls on the new
RSHF line when bypassing return air only.
Bypassing a mixture of outdoor and return air
causes the mixture point (3) to fall on the GSHF
line, I;ig. 60. The air is then supplied to the space
along the new RSHF line (not shown in Fig. 60) at
a higher moisture content than the air supplied
when bypassing return air only. Thus it can be
readily observed that humidity control is further
hindered with the introduction of unconditioned
outdoor air into the space.
VOLUME CONTROL
Volume control of the supply air quantity provides essentially the same type of control that results
I’AKT I. LOAD ES’I‘IMATING
1-144
Erom bypassing return air around the heat transfer
equipment, I;ig. 60. However, this type of control
may produce problems in air distribution within
the space and, therefore, the required air quantity
at partial load sl~ould be evaluated for proper air
distribution.
- R E T U R N AIR
CONDITIONED
SPACE
@
c
’
OUTDOOR
AIR
-
/ BYPASS
AIR
4
F
-
FAN
APPARATUS
2
3
&-
FIG. 61- SCHEMATIC SKETCH OF B YPASS CONTROL
W ITH BYPASS OF RETURN A IR O NLY
ON-OFF CONTROL OF AIR HANDLING EQUIPMENT
On-off control oE air handling equipment (Eancoil units) results in a fluctuating room temperature
and space relative humidity. During the “off” operation the ventilation air supply is shut off, but
chilled water continues to flow thru the coils. This
method of control is not recommended for high
latent load applications, as control of humidity may
be lost at reduced room sensible loads.
.-
ON-OFF CONTROL OF REFRIGERATION EQUIPMENT
On-off control of refrigeration equipment (large
packaged equipment) results in a fluctuating room
temperature and space relative humidity. During
the “oW operation air is available for ventiIation
purposes but the coil does not provide cooling.
Thus, any outdoor air in the system is introduced
into the space unconditioned. Also the condensed
moisture that remains on the cooling coil, when the
reErigeration equipment is turned ofE, is re-evaporated in the warm air stream. This is known as
re-evaporation. Both oE these conditions increase
the space latent load, and excessive humidity results. This method of control is not rccommendcd
for high latent load applications since control of
humidity may be lost at decreased room sensible
loads.
REFRIGERATION CAPACITY CONTROL
Relrigeration capacity control may be used on
either chilled water or direct expansion ret‘rigeration equipment. Partial load control is accomplished on chilled water equipment by bypassing
the chilled water around the air side equipment
(Ean-coil units). Direct expansion refrigeration
equipment is controlled either by unloading the
compressor cylinders or by back pressure regulation ’
in the refrigerant suction line.
Refrigeration capacity control is normally used
in combination with bypass or reheat control. When
used in combination, resuIts are excellent. When
used alone, results are not as effective. For example,
temperature can be maintained reasonably well, but
relative humidity will rise above design- at partial
load conditions, because the latent load may not
reduce in proportion to the sensible load. *
PARTIAL’ LOAD CONTROL
Generally, reheat control is more expensive but
provides the best control of conditions in the space.
Bypass control, volume control and refrigeration
capacity control provide reasonably good humidity
control in average or high sensible heat factor
applications, and poor humidity control in low
sensible heat Eactor
applications. On-off control
usually results in the least desirable method of
maintaining space conditions. However, this type
of control is frequently used for high sensible heat
factor
applications with reasonably satisfactory
results.
CH,\I”I‘I-K
8 . /\I’I’I,IEI~
I’SY~:FI11OILlE’I‘I~I~:S
l-145
TABLE 65-APPARATUSDEWPOINTS
EFFECTIVE SENSIBLE HEAT FACTOR
AND
APPARATUS
DEWPOINT”
9 0 - 80 F DB
ROOM
CONDITIONS
DBlRHl WBI W
EFFECTIVE SENSIBLE HEAT FACTOR
AND APPARATUS DEWPOINT*
(F) I / (%) / (F) // (fir/lb)
45
bb7.
73S. ESHF
ADP
1 . 0 0 .91 .87 .80 .75 .72 .bB’.bS .43
58.5 57 56 54 52 50 46 41 53
Ss
b9. 8
9.. 2 ESHF
ADP
1 . 0 0 .90 .83 .74 .bB .b4 . b l .58 .5b
6 4 . 2 63 62 60 58 56 54 50 44
bs 72. 8 ,07 . o
I ,
,4 9 ,ob 4 ESHF
.
.
ADP
1 . 0 0 .92
60.9 68
.?8
66
.b4 .bO S8
6 1 5 8 56
63
Sb .54
53 47
55 7b . 7 ,,,. S ESHF 1 . 0 0 .92 .7b .bB .b4 .57 .54 .52 .50
ADP 7 1 . 6 71 69 67 66 62 59 57 50
,
b. 78 . 4 128 . 4 ESHF 1 . 0 0 .Bb .bB .bO Sb .52 .SO .48 .4b
ADP 7 4 . 2 73 71 69 67 64 62 59 50
bs
8.. o 139 . b ESHF
ADP
1 . 0 0 .75 .bB .b2 .SS .SO .47 .45 ,43
7 6 . 8 75 74 73 71 69 66 64 59
7. 81 . b lsl . o ESHF
ADP
1 . 0 0 .78 .$b .bO .S2 .47 .43 .41 39
7 9 . 0 78 7 7 , 7 6 74 72 69 66 58
/
I
ESHF
ADP
1 . 0 0 .Bb .71 .b3 .58 .54 .S2 .Sl .49
69.1 68 66 64 62 60 58 56 51
7o
74. 2 ,ls. s ESHF
ADP
1 . 0 0 .80 .71 .bS .bO .54 Sl .48 .4b
7 1 . 2 7 0 69 68 67 65 63 60 56
3s
b25 .
Ss2 . ESHF
ADP
1 . 0 0 .94 .89 .84 .81 .77 .75 .73 .71
5 0 . 8 49 47 45 43 39 36 32 21
4. b4 . 2
b3. 2 ESHF
ADP
1 . 0 0 .94 .87 .02 .78 .75 .72 .b9 .b7
5 4 . 4 53 51 49 47 45 41 36 23
4S bs9.
71 . 2 ESHF
ADP
1 . 0 0 .9b .91 .83 .78 .74 .70 .b7 .b4
5 7 . 6 57 56 54 52 50 47 43 36
I
81
ss
b9 o 874 E S H F 1 . 0 0 .90 .77 .71 . b b .b2 . b O .58 .56’
.
.
ADP 63.2 62 60 58 56 53 51 47 35
b.
7.. s
bS
9s . 4 ESHF
ADP
71 . 9 ,037. ESHF
ADP
I
I
I
7.. 8
90,. ESHF
ADP
” 72-3i 99-4
ESHF
ADP
1 . 0 0 .85 .76 .71 .bb .bO .Sb .S2 SO
6 8 . 2 6 7 66 65 64 62 60 56 52
I
7. 73 . 3 ,,, . 9 ESHF
ADP
so
1 . 0 0 .92 .77 .68 .63 .59 Sb .54 .53’
6 5 . 8 65 63 61 59 56 53 50 4 6
/
I
I
I
I
/
I
I
1 . 0 0 .80 .71 .bl SS .52 .48 .47 .46
7 0 . 3 69 68 66 64 62 58 56 52
1 . 0 0 .92 .80 .73 .bB .b4 . b l .S9 .57
6 4 . 2 63 61 59 57 54 51 40 39
1 . 0 0 .92 .83 .73 .b7 .bO .S7 .Sb .54
6 6 . 9 66 65 63 61 57 54 52 4 7
bs 7s . S 1182 ESHF 1 . 0 0 .8B1 .b9 .bl .Sb .53 .SO .48 .47
.
ADP 7 1 . 9 71 69 67 65 63 61 58 54
/
7. 77 . o ,27. b ESHF/ 1 . 0 0 .81 .b3 .SS
ADP ! 7 4 . 0 73 71 69
82 _
3S b33 .
570 ESHF
.
ADP
1 . 0 0 .92 .88 .84 .80 .7b .74 .72 .71
5 1 . 6 49 48 4 6 4 3 3 9 3 6 31 27
4o
“*’ ESHF
ADP
1 . 0 0 .90 .87 .82 .78 .74 .71 .b9 .b7
5 5 . 2 53 52 50 48 45 41 38 31
“*’
See page 147 for notes.
ESHF
ADP
1 . 0 0 .78 .71 .bS . b l .SS .52 .49 .476 9 . 4 68 6 7 , 66 65 63 61 58 5 3 ,
I
1-146
PART
TABLE
79 - 72 F DB
65-APPARATUS DEWPOINTS (Continued)
ROOM
CONDITIONS
EFFECTIVE SENSIBLE HEAT FACTOR
AND
APPARATUS
DEWPOINT*
35
5e’9
EFFECTIVE SENSIBLE HEAT FACTOR
A N D A P P A R A T U S DEWPOlNT*
46-7
1 . 0 0 .96 .91 .87 .83 .79 .77 .75 .73
40.2 4 7 4 5 4 3 41 3 7 3 5 31 2 2
4. 6, . 9
573 . ESHF
ADP
1 . 0 0 .93 ,07 .82
5 1 . 7 50 4 8 46
1
I
1
5’ 65’o
78
s5
I
66 . 6
I
I
I
I
I
I
I
I
1 . 0 0 .94 .06 .81 .77 .74 .71 .69 .67
53.2
52 50 48 4 6 4 4 4 0 3 7 31
67’4 ESHF
ADP
1 . 0 0 .93 ‘.83 .77 .73 .69 .67 .65, .63
56.2 5 5 5 3 5 1 4 9 46 4 3 4 0 3 2
693 .
64’9
74’o
ESHF
ADP
1 . 0 0 .94 .82 .75 .70 .67
5 8 . 7 58 56 5 4 5 2 50
6o
66’2
ao’9
ESHF
ADP
1 . 0 0 .90 .77 .70 .66 .62 .60 .58 .57
61.1 6 0 5 8 56 5 4 5 2 4 9 4 6 4 3
65
67 . 6
87 . 6 ESHF
ADP
1 . 0 0 .84 .72 .65 .61 .58 .56 .54 .53
63.4 6 2 6 0 5 8 5 6 5 4 5 2 4 8 4 3
TO
68’9
94’6 ESHF
ADP
1 . 0 0 .a0 .67 .60 .56 .54 .52 .51 .50
65.5 64 6 2 6 0 5 8 56 5 4 5 2 4 9
!
.64 .62
71’9 ESHFADP 57.9 1.00 .94 57.03 55.76 53.73
51 .70 49.67 47 42 36
!
6. 67 9 86 4 ESHF 1.00 .90 .82 .76 .69 .64 .60 .57 -55
.
.
ADP 63.0 62 61 60 58 56 53 49 42
65
55
!
79 . 2 ESHF 1 . 0 0 ,96 .83 .75 .70 .65 .62 .60 .59
ADP 60.5 6 0 5 8 5 6 5 4 51 48 4 4 41.
I
I
I
/
/
I
/
I
/
r
93 . 8 ESHF
ADP
1 . 0 0 .85 .77 .71 .67 .62 .58 .54 .52
6 4 6 3 6 2 61 5 9 5 7 5 3 4 8
7. 7.. 6 ,o, . 2 ESHF
ADP
1 . 0 0 .71 .66 .62 .59 .55 .52 .50 .48
67.5 6 5 6 4 6 3 62 6 0 5 8 5 5 4 8
75
7. 68 o 91 2
.
.
601 6 7 . 1 1 83.61
70
69.8
97.
1 . 0 0 .96 .89 .84 .81 .78 .76 .72 .70
4 9 . 9 49 47 45 43 41 39 32 22
604 . ESHF
ADP
6, . 9
.79 .77 .73 .71 .69
44 42 38 34 25
I
ESHF
ADP
45
I
5oo . ESHF
ADP
1 . 0 0 .96 .91 . a 7 .84 .8i .79 .77 .74
46.3 4 5 4 3 41 39 37 3 4 31 21
537 .
76
6.f3
ESHF
ADP
4. 604 .
c:-’ 63’4
35
I. LOAD ESTIMATING
.65 .62 .60
48 44 38
,
~H,\I’~I‘EII
1-147
H. ;\l’l’l.lEI) I’SY(:HKOMETRICS
TABLE
CONDITIONS
65-APPARATUS DEWPOINTS (Continued)
ROOM
CONDITIONS
EFFECTIVE SENSIBLE HEAT FACTOR
A N D A P P A R A T U S DEWPOINT*
DB/RH WB/
(F)
(F) I(%) 1(F) 1(er/lb)
” 64.0
76.3
7o
82.3
72
65.2
ESHF
ADP
(%I (F)
1 . 0 0 .84 .73 .67 .63 .61 .59 .S8
5 9 . 5 58 56 54 52 50 48 47
ESHF 1 . 0 0 .80 .69 .62 .59 .56 .54 .53
ADP 161.61 6 0 58 56 54 51 48 44
W
7 2 - 5 5 F DB
EFFECTIVE SENSIBLE HEAT FACTOR
A N D A P P A R A T U S DEWPOINT+
(w/lb)
6. S23 .
E S H F ’ l . O O .94 .89 .81 .77 .74 .72 .70 .68
462 . ADP 46.0 45 44 42 40 38 36 34 28
”
50*o
53*3
I70 114.31
53.91 ESHF
ADP / 50.1 1 * 00 i1. 49 89 1. 48 83 1. 74 46 1. 701.67 44 142
I
75
E S H F 1 . 0 0 .91 .86 .78 .74!.70 .69 .67 .65
A D P 4 8 . 1 47 4 6 , 44 421 4 0 39 36 31
I
55’3
57.8
I
ESHF
ADP
I
I
I
I
1.001.79I.74 .71 .68 .64 .62 .60 .59
5 2 . 0 / 50 14 9 48 47 45 43 / 40 37
*O 56+3
“e7 ESHF
ADP
1 . 0 0 .85 .76 I.70 .66 .61 .59 .57 .56
5 3 . 8 i 5 3 , 52 51 50 48 46 44 41
9o
58’2
69*4
ESHF
ADP
95
59-1
73*5
E S H F 1 . 0 0 / .69 1 . 5 5 .49 .47 .46 .45
ADP
58.5 / 58 57 56 55 54 52
1 . 0 0 .72 .62 .57 .54 .52 S O .49
5 7 . 0 56 55 54 53 52 50 47
-
T
70
55
i
ES
9o
ES
66 . 8
937. ESHF
ADP
1 . 0 0 .71 .56 .52 .50 .48 .47 .46 .45
6 5 . 3 64 62 61 60 59 58 57 54
9. 67 . 9
99 . 3 ESHF
ADP
1 . 0 0 .66 .56 .50 .47 .45 .43 .42 .41
6 6 . 9 66 65 64 63 62 61 60 56
ESHF
5’*2 ADP
*O 5’-5
S24 . S4 . S ESHF
53*2
s7*7
1 . 0 0 .88 .79 .74 .67 .64 .62 .6l .60
48.8 48 47 46 44 42 40 39 37;
ADP
1 . 0 0 .77 .70 .66 .63 .60 .58 .57
5 0 . 4 49 48 47 46 44 42 40
ESHF
ADP
1 . 0 0 .76 .67 .61 SE .55 .54 .53
5 2 . 0 51 50 49 48 46 44 41
i
* T h e v a l u e s shown i n t h e g r a y areas i n d i c a t e t h e l o w e s t e f f e c t i v e
sensible heat factor possible without the use of reheat. This limiting
condition is the lowest effective sensible heat factor line that intersects the saturation cuwc. Note that the r~~rn dewpoint is equal to
the required apparatus dewpoint for on effective sensible heat factor
of 1.0.
6s
65
57 . 7
59 7 ESHF
.
ADP
1 . 0 0 .92 .85 .80 .73 .69 .66 .64 .62
5 2 . 9 52 51 50 48 46 44 41 37
7. SE . 9
64 . S ESHF 1 . 0 0 .89 .80 .76 .69 .65 .62 .60’.58
ADP 5 5 . 0 54 53 52 50 48 46 4 3 3 7
75
S9 . 9
69 . 2 ESHF
ADP
1 . 0 0 .88 .78 .72 .651.61 .58 .56 .55
5 6 . 9 56 55 54 5 2 5 0 48 45 41
8. S, . o
73 . 8 ESHF
ADP
1 . 0 0 .75 .68 .63
5 8 . 7 57 56 55
/
620.
I
83 . 2 ESHF
ADP
1 . 0 0 .70 .58 .53 .50 .48 .46 .45
6 1 . 9 61 60 59 58 57 55 53
95
880 . ESHF
ADP
1 . 0 0 .69 .51 .46 .43 .42 .41
6 3 . 5 63 62 61 60 59 58
=
1
, +.62a (wm- wad,)
(trm - tad,,)
786 . ESHF 1 . 0 0 .71 .63 .58 .55 .52 .50 .49
ADP 6 0 . 3 59 58 57 56 54 52 50
9. 63 . o
640 .
1 . F o r R o o m C o n d i t i o n s N o t G i v e n ; T h e a p p a r a t u s dewpoint m a y
be determined from the scale on the chart, or may be calculated
CIS shown in the following equation:
ESHF
/
/
ES
NOTES FOR TABLE 65:
This equation in mwe familiar form il:
E S H F
0 . 2 4 4 (frm - t.dp)
= - 0 . 2 4 4 (trm -todp) + g (wr, - W.dp)
(Cont.)
1-148
i
1’.4KT
I. LOAD ESTIMATING
7000 = groins per pound.
w h e r e W,, = room moisture content, gr/lb of dry air
W.dp
= moisture content at apparatus dewpoint,
gr/lb of dry air
2.
frm
= room dry-bulb temperature
todp
=apporatus
3 . F o r A p p a r a t u s Dewpoint B e l o w F r e e z i n g . T h e l a t e n t h e a t o f
fusion of the moisture removed is not included in the calculation of
apparatus dewpoint below freezing or in the calculation of room
load, in order to simplify estimating procedures. Use the some
equation as in Note 1. The selection of equipment on CI basis of
16 to 1 B hour operating time provides a safety factor large enough
to cover the omission of this latent heat of fusion, which is a small
part of the total load.
dewpoint
temperature
0.244 = specific heat of moist air at 55 F dewpoint,
Btu per deg F per lb of dry air
1076 = average heat removal required to condense
one pound of water vapor from the room air
For High Elevations. For
elevations, see Table 66.
effective
sensible
heat
factors
at
high
TABLE 66-EQUIVALENT EFFECTIVE SENSIBLE HEAT FACTORS FOR VARIOUS ELEVATIONS*
For use with sea level psychrometric chart or tables
Effective
Sensible Heat
Factor from Air
1000
Conditioning
(28.86)
Lo(ld E s t i m a t e -
.60
.55
.50
Elevation
2000
(27.82)
Equivalent
.61
.56
.51
.62
.57
.52
(Feet)
3000
(26.82)
Effective
Sensible
.63
.58
.53
end
Barometric
4000
(25.84)
Heat
5000
(24.89)
Factor
.64
.59
.54
Referred
.6A
.60
55
1
ESHF== IpI) (I-ESHF) +,
(pc.1 W-IF)
= barometric pressure at sea level
= barometric pressure at high elevation
PI
ESHF = ESHF obtained from air conditioning load estimate
ESHF,=
5TES
FOR
equivalent ESHF referred to a sea
psychrometric chart or Table 66
TABLE
level
66:
1. The required apparatus dewpoint for the high elevation is determined from the sea level chart or Table 65 .by use of the equivalent
effective sensible heat factor. The relative humidity and dry-bulb
temperature must be used to define the room condition when using
this table because the above equation was derived on this basis.
The room wet-bulb temperature must not be used because the wet-
(Inches
6000
(23.98)
to o
of
Hg)
7000
(23.09)
et
Installation
8000
(22.12)
Sea Level Psychrometric
.65
.61
.56
.66
.61
.57
.67
.62
.57
.
9000
(21.39)
Chart or
10000
(20.57)
Tables
.68
.63
.5B
-
.69
.64
.59
bulb temperature corresponding to any particular condition, for
example, 75 F db, 40% rh, at o high elevation is lower (except
for saturation) than that corresponding to the same condition (75 F
db, 4Oa/o rh) at sea level. For the same value of room relative
humidity and dry-bulb temperature, and the some apparatus dewpoint, there is a greater difference in moisture content between
the two conditions at high elevation than at sea level. Therefore,
o higher apparatus dewpoint is required at high elevation for a
given effective sensible heat factor.
*Values obtained by use of equation
W h e r e p0
Pressure
2. Air conditioning load estimate (See Fig. 44). The fack 1.08 ond
.68 on the air conditioning load estimate should be multiplied by
(Pl)
the direct ratio of the barometric pressures -. Using this method,
(P01
it is assumed that the air quantity (cfm) is measured at actual conditions rather than at standard air conditions. The outdoor and
room moisture contents, grains per pound, must also be corrected
for high elevations.
3.
Reheat-Where the equivalent effective sensible heat factor
lower than the shaded values in Table 65, reheat is required.
is
<;HAITER 8 . XI’I’LIED
1-149
I’SYCHKOMETRICS
SYMBOLS
ABBREVIATIONS
adp
apparatus tlcwpoint
r5F
(IZF) (OALH)
(RF) (OASH)
(RF) (OATH)
lStu/hr
bypass factor
bypassed outdoor air latent heat
bypassed outdoor air sensible heat
bypassed outdoor air total heat
british thermal units per hour
crm
cubic feet per minute
(lb
dp
dry-bulb
dcwpoint
ERLH
ERSH
ERTH
ESHF
effective
effective
effective
effective
F
fprn
Fahrenheit degrees
feet per minute
SP
9-P
GSHF
GTH
GTHS
gallons per minute
grains per pound
grand sensible heat factor
grand total heat
grand total heat supplement
OALH
OASH
OATH
outdoor air latent heat
outdoor air sensible heat
outdoor air total heat
rh
RLH
RLHS
RSH
RSHF
RSHS
R’
relative humidity
room latent heat
room latent heat supplement
room sensible heat
room sensible heat factor
room sensible heat supplement
room total heat
Sat Elf
SHF
saturation efficiency of sprays
sensible heat factor
TLH
TSH
total latent heat
total sensible heat
wb
wet-bulb
In ba
cim da
ch,,,
cfm,.,
c/m,,
bypassed air quantity around apparatus
dehumidified air quantity
outdoor air quantity
return air quantity
supply air quantity
specific enthalpy
apparatus dcwpoint enthalpy
effective surface temperature enthalpy
entering air enthalpy
lcaving air enthalpy
mixture of outdoor and return air
cnthalpy
outdoor air enthalpy
room air enthalpy
supply air enthalpy
room latent heat
room sensible heat
room total heat
sensible heat factor
t
t Wfp
t rffh
t.
L
L~h
tldb
t Ito
t itch
t 1,L
t Ml
t,.,,,
tm
W
Wodp
WC,
We,
WI,
W,
W,*
W,?ll
W,,
temperature
apparatus dewpoint temperature
entering dry-bulb temperature
effective surface temperature
entering water temperature
entering wet-bulb temperature
leaving dry-bulb temperature
leaving water temperature
leaving wet-bulb temperature
mixture of outdoor and return air
dry-bulb temperature
outdoor air dry-bulb temperature
room dry-bulb temperature
supply air dry-bulb temperature
moisture content or specific humidity
apparatus dewpoint moisture content
entering air moisture content
effective surface temperature moisture
content
leaving air moisture content
mixtureof outdoor and return air
moisture content
outdoor air moisture content
room moisture content
supply air moisture content
l-150
PART 1. LOAD ESTIMATING
PSYCHROMETRIC
A.
FORMULAS
AIR MIXING EQUATIONS (Outdoor and Return Air)
h, = (cfm”a x hoa) + (cfm,, x km)
Cf%a
w n = (cfm,, x WA + (cfs x WA
cfm,,
(2)
TSH
TLH
GTH
(3)
RSHF =
ESHF =
ERSH = RSH + (BF) (OASH) + RSHS* (4)
ERLH = RLH + (BF) (OALH) + RLHS’ (5)
ERTH = ERLH + ERSH
(6)
TSH
TLH
GTH
= RSH + OASH + RSHS”
= RLH + OALH + RLHS”
= TSH + TLH + GTHS”
RSH
RLH
RTH
RTH
=
1.08-f
=
.68t
x cfm, x (tr,,, - t,,)
x cfm,, x (W,, - W,,)
4.45-t x cfm,, x (h,, - haa)
= RSH $ RLH
OASH = 1.08 x cfm,, (t,, - t,,)
O A L H = .68 x cfm,, ( W , , - W,,)
OATH = 4.45 x cfm,, (h,, - h,)
(7)
(8)
c-9
(14)
(15)
(16)
OATH = OASH + OALH
(17)
(BF) (OATH) = (BF) (OASH) + (BF) (OALH)
(18)
ERSH = 1.08 x
ERLH
=
.68 x
ERTH = 4.45
x
cfm& x (t, - tadp) (1 - BF)
cfq,J x
cfmaa$
X
(19)
(W,, - Wadp) (1 - BF)
(20)
(h, - ha& (1 - BF)
(21)
*RSHS, RLHS and GTHS are supplementary loads due to
duct heat gain, duct leakage loss, fan and pump horsepower
gains, etc. To simplify the various examples, these
plementary loads have not been used in the calculations.
However, in actual practice, thesesupplementary loads should
be used where appropriate. Chapter 7 gives the values for the
various supplementary loads. Fig. 1, Chapter I, illustrates the
method of accounting for these supplementary loads on the
air conditioning load estimate.
fItem H, page 151, gives the derivation of these air constants.
fWhen no air is to he physically bypassed around the conditioning apparatus, cfnd, = cfm,,.
.
RSH
RSH
R S H + R L H =--RTH
ERSH
D.
ERSH
ERTH
ERSH + ERLH = ~
GSHF =
(23)
(24)
TSH
=GTH
TSH
TSHfTLH
(25)
(26)
(27)
.
BYPASS FACTOR EQUATIONS
t edb
BF = hdb - todp ; (1 _ BF) =
(10)
(11)
(12)
(13)
(22)
C. SENSIBLE HEAT FACTOR EQUATIONS
B. COOLING LOAD EQUATIONS
=
= 1.08 X cfm& X (tcdb - tJdt,)“*
= .68 x kfmaaz X (W,, - W,,)“*
= 4.45 x cfm& x (hoa - h,,)**
t edb - bdp
- hdb
t edb - bdp
(28)
BF = wl, - Wadp ; (1 - RF) = pI;za
we, - Wadp
ea
a&
(29)
.
BF = 4, - hadp ; (1 - BF) = ,“eaI,“ia
(30)
ea
:
he, - hodp
a&
E. TEMPERATURE EQUATIONS AT APPARATUS
x toa) + (cfm, x LJ
t edb
tZdb
(31)
cfm,,f
= bdp + BF (kdb - kdp)
t ewb and hzob
correspond
to
the
(32)
calculated
values
of h,, and hl, on the psychrometric chart.
h ea xI
= (cfm,, X hoa) + (cfm, x h,)
(33)
cf md
h la
= hadp + BF (he, - had,)
(34)
F. TEMPERATURE EQUATIONS FOR SUPPLY AIR
t,, = bn -
RSH
1.08 (cfm,,U
(35)
**When t,,,, W,,, and hm are equal to the entering conditions
at the cooling apparatus, they may be substituted for tedb.
W,, and hea respectively.
(;t-lAI”I‘El< x.
/\l’l’I.IIcl)
I’SY(:HKOIZlE~I’III~:S
1-151
G. AIR QUANTITY EQUATIONS
\_
_
Cf%,z
ERSH
= 1.08 x (1 - BF) (trm - t,,,)
cfmcta = ~
ERLH
~33 x (I - RF) (JJ’,,,, - Wertp)
cf%in
ERTH
=
4.45 x (1 - BF) (A,, - h,,J
cfmd =
cfmdd =
cffn&
=
TSH
I.08
(t,,b - t,rid
TLH
.@3 (We, - W,“)
GTH
4.45 (II,.,‘ - h,,,)
drn bn
= cf7n,,
Note:
cftnd,,
(,.flfl ,“R
= ffm,,,, + cfttt,.,,
(36)
(37)
H.
Cf%,, =
\
x (t,.,,, - Lzl
.x’ .,’ ,’ )
‘1.
-. ~.-RLH- -.. /-‘Cfms,L =
.68 x (W,.,,, - W,,,)
cfm,,
=
1.08
(38)
RTH
4.45 x (h,.,,‘ - h,,)
where
.244 = specific heat of moist air at
70 F db ant1 50% rh,
Btu/(deg F) (lb dry air)
60 = min/hr
13.5 = specific volume of moist ail
at 70 F c!b and 50% rh
.&3
Ix 60-
x -
13.5
where
(41)
(42)
(44)
Ct*G)
.244 x 60
13.5
(39)
(43)
(‘9
will bc less than cfrn,, only when
air is physically bypassed around the
conditioning apparatus.
,/.’ ,------a
<
cfmd,,
DERIVATION OF AIR CONSTANTS
1.08 =
(40)
-
4.45 =K
13.5
where
1076
7000
60 = min/hr
13.5 = specific volume of moist air
at 70 F db and 50% rh
1076 = average heat removal required to condense one
pound of water vapor from
the room air
7000 = grains per pound
60 = min/hr
13.5 = specific volume of moist air
at 70 F db and 50% rh
l-152
BIBLIOGRAPHY
CHAPTER 2
5.
Air Conditioning and Refrigeration Institute, Application
Engineering Standard 530-56 Air Conditioning. 1956.
2. Heating, Ventilating and Air Conditioning Guide, Chapters 12 and 13, 1956.
3. Summer Weather Data - Marley Company.
6.
TABLE 1
1.
TABLE 4
1. Conditions for Comfort, by C . S. Leopold: H e a t i n g ,
Piping and Air Conditioning, June 1947, p. 117.
2. The Mechanism of Heat Loss and Temperature Regulation, by E. F. DuBois; Transactions of the Association of
American Physicians, Vol. 51, 1936, p. 252.
3. The Relative Influence of Radiation and Convection upon
the Temperature Regulation of the Clothed Body, by
C. E. A. Winslow, L. P. Herrington, and A. P. Gagge;
American Journal of Physiology, Vol. 124, October 1938,
p. 51.
4. Reactions of Office Workers to Air Conditioning in South
Texas, by A. J. Rummel, F. E. Giesecke, W. H. Badgett,
and A. T. Moses; ASHVE Trans., Vol. 45, 1939, p. 459.
5. Shock Experiences of 275 Workers After Entering and
Leaving Cooled and Air Conditioned Offices, by A. B.
Newton, F. C. Houghten, C. Gutherlet, R. W. Qualley,
and M. C. W. Tomlinson; ASHVE Trans., Vol. 44, 1938,
p. 571.
6. How to Use the Effective Temperature Index and Comfort Charts, by C. P. Yaglou, W. H. Carrier, Dr. E. V. Hill,
F. C. Houghten, and J. H. Walker; ASHVE Trans., 1932,
p. 411.
7. Heat and IMoisture Losses from the Human Body and
Their Relation to Air Conditioning Problems, by F. C.
Houghten, W. W. Teague, W. E. Miller and W. P. Yant;
ASHVE Trans., 1929, p. 245.
8. Thermal Exchanges Between the Human Body and Its
Atmospheric Environment, by F. C. Houghten, W. W.
Teague, W. E. Miller and W. P. Yant; American J o ~ w n a l
of Physiology, Vol. 88, 1929, p. 386.
TABLE 5
1. Heating, Ventilating and Air Conditioning Engineers
Guide, Chapter 45, 1956.
2. Air Conditioning and Refrigerating Data Rook, 1955.
7.
8.
9.
10.
Circuit i\naIysis Applied to Load Estimating, Part IIInfluence of Transmitted Solar Radiation, by H. B. Nottagc and C. V. Parmelee; ASHAE Transactions, Vol. 61,
1955, pp. 128-139.
Thermal Circuit Analysis for Developing Application
Engineering Information, hy Stanley F. Gilman
and 0.
\Villiam Clausen; Ifeating, Piping and Air Conditioning,
June 1957, pp. 153-160.
Temperature Changes in Refrigerated Rooms During
I’ulldown P e r i o d , hy J. L. Threlkeld and T. Kusada;
Refrigerating Engineering, July 1956, p. 35.
Heat Transmission as influenced by Heat Capacity and
Solar Radiation by F. C. Houghton, J. L. Blackshaw, E.
M. Pugh, and P. McDemott; ASHAE Transactions, Vol. 38,
1932. p. 263.
Cooling Load from Sunlit Glass, hy C. 0. Mackey and
N. R. Gay: ASHAE Transactions, Vol. 58, 1952, p. 321.
Analysis of an Air Conditioning Thermal Circuit by an,
Electronic Differential Analyzer, by G. V. Parmelee, P.
Vance, and A. N. Cerny; Heating, Piping and Air Conditioning, Sept. 1956, p. 117.
TABLE 15
1. The Transmission of Solar Radiation Through Flat
Glass Under Summer Conditions, by G. V. Parmelee;
Heating, Piping and Air Conditioning, October-November 1945. Also ASHVE Trans., 1945, Vol. 51, p. 317.
2. Measurements of Solar Radiation Intensity and Determinations of its Depletion by the Atmosphere,aby
H. H.
Kimball; Monthly Weather Review, February 1930, Vol.
58, p. 52.
3. Review of United States Weather Bureau Solar Radiat i o n I n v e s t i g a t i o n s , b y I . F . H a n d ; Monthly Weather
Review, December 1937, Vol. 65, p. 430.
4.
Smithsonian
p. LXXXIV.
Meteorological
Tables,
5th
Revised
edition,
5. Pyrheliometers and Pyrheliometric Measurements, by
I. F. Hand; U. S. Weather Bureau, November 1946.
6. Proposed Standard Solar Radiation Curves for Engineering Use, by Parry Moon; Journal of the Fra nklin
Institute, November 1940, Vol. 230, No. 5, pp. 586-617.
7. Performance of Flat Plate Solar Heat Collectors, by
H. C. Hottel and B. B. Woertz; ASME Trans., February
1942, Vol. 64, pp. 91-104.
8. Where is the Sun?, by M. J. Wilson and J. M. Van
Swaay; Heating and Ventilating, May and June 1942.
9.
CHAPTER 3
TABLES 6 THRU 12
I. Heat Transfer, hy Max Jakoh, Vol. 1, John Wiley 9c Sons,
Inc., New York, N. Y., 1949.
2. The Solution of Transient Heat Conduction Problems by
Finite Differences, by G. A. Hawkins and J. T. Agnew.
Purdue Univ., Eng. Exp. Sta. Bulletin No. 98, 1946.
3. Hydraulic Analogue for the Solution of Problems of
Thermal Storage, Radiation, Convection and Conduction,
by C. S. Leopold; ASHAE Transactions, Vol. 54, 1948, p.
389.
4. Circuit Analysis Applied to Load Estimating, by H. B.
Nottage and G. V. Parmelee; ASHAE Transactions, Vol.
60, 1954, pp. 59-102.
.
Summer Weather Data and Sol-Air Temperature Study of Data for New York City, by C. 0. Mackey and
E. B. Watson: Heating, Piping and Air Conditioning,
Nov. 1944, p. 651. Also ASHVE Trans., 1945, Vol. 51,
p. 75.
10. Summer Weather Data and Sol-Air Temperature Study of Data for Lincoln, Nebraska, by C. 0. Mackey
and E. B. Watson; Heating, Piping and Air Conditioning, January 1945, p. 42. Also ASHVE Trans., 1945, Vol.
51, p. 93.
TABLE 16
1. An Experimental S t u d y of Flat-type Sun Shades, by
G. V. Parmalee, W. W. Aubele and D. J. Vild; Heating,
Piping and Air Conditioning, January 1953.
BIBLIOGRAPHY
l-153
2. Design Data for Slat-type Sun Shades for Use in Load
Estimating, by G. V. Pnrmalee and D. J. Vild; Heating,
Piping and Air Conditioning, September 1953.
3. The Transmission of Solar Radiation Through Flat
G l a s s u n d e r S u m m e r C o n d i t i o n s , b y G . V . Parmalee;
Heating, Piping and Air Conditioning, October, November 1945. Also ASHVE Trans., 1945, Vol. 51, p. 317.
4.
Heat Gain Through Western Windows With and Without Shading, by F. C. Houghten and David Shore;
Heating, Piping and Air Conditioning, April 1941, p,
256. Also ASHVE Tmns., 1941, pp. 251.274.
5. Studies of Solar Radiation Through Bare and Shaded
Windows, by F. C. Houghten, C. Gutberlet, and J. Blackshaw; ASHVE Trans., 1934, Vol. 40, pp. 101-116.
6. Solar Heat Gain Factors For Windows With Drapes, by
N. Ozisik and L. F. Schutrum; paper presented at
ASHR,lE
meeting, Dallas, Texas, February, l-4, 1960.
TABLES 35 AND 36
I.
Houghten, S. I. Taimuty, C. Gutberlet and C. J. Brown;
ASN VE Trans., Vol. 48, 1942, p. 369.
2. Measurements of Heat Losses From Slab Floors, by R. S.
Dill, W. C. Robinson and H. E. Robinson; U. S. National
Bureau of Standards, Report BMSIOJ, 1945.
3. Heat Losses Through Floors of Basementless Building;
Heating and Ventilating, Vol. 42, Septemher 1945, p. 89.
TABLE 38
1. Ice Formation on Pipe Surfaces by S. Lewis Elmer, Jr.;
Re/rigerating
Engineering, July 1332, p. 17.
2. Notes on the Formation of Ice on Pipe Surfaces by F.
Raseri; Refrigerating Engineering, January 1933, p. 21.
TABLE 40
TABLE 17
1. Heat Gain Through Glass Blocks by Solar Radiation
and Transmittance, by F. C. Houghten, David Shore,
H. J. Olson and Burt Gunst; ASHVE Trans., 1940, pp.
83-107:
1.
2.
CHART 1 AND TABLE 18
1 Tables of Computed Altitude and Azimuth, Volume I
to V inclusive (0” to 50” Latitude, 10” per volume) ;
U. S. Navy Department - Hydrographic Office, No. 214.
2. Where is the Sun?, b y M . J. W i l s o n a n d J . M . V a n
Swaay; Heating and Ventilating, May, June 1942.
3.
4.
CHAPTER 5
TABLES 19 AND 20
1. Periodic Heat Flow - Homogeneous Walls or Roofs, by
C. 0. Mackey and L. T. Wright, Jr.; Heating, Piping and
Air Conditioning, September 1944, p. 546. Also ASHVE
Trans., 1944, Vol. 50, p. 293.
5.
2. Periodic Heat Flow - Composite Walls or Roofs, by C. 0.
Mackey and L. T. Wright, Jr.; Heating, Piping and Air
Conditioning, June 1946, p. 107.
6.
3. Summer Cooling Load as Affected by Heat Gain Through
Dry, Sprinkled. and Water Covered Roofs, by F. C.
Houghten, H. T. Olson, and Carl Gutberlet; Heating,
?ping and Air Conditioning, July 1940. Also ASHVE
:rans., 1940, p. 231.
4. Summer Weather Data and Sol-Air Temperature - Study
of Data for New York City, by C. 0. Mackey and E. B.
Watson; Heating, Piping and Air Conditioning, November 1944, p. 651. Also ASHVE Trans., 1945, Vol. 51, p. 75.
5. Summer Weather Data and Sol-Air Temperature - Study
of Data for Lincoln, Nebraska, by C. 0. Mackey and E. B.
Watson; Heating, Piping and Air Conditioning, January
1945, p. 42. Also ASHVE Trans., 1945, Vol. 51, p. 93.
6. Estimating Heat Flow Through Sunlit Walls, by C. 0.
Mackey and L. T. Wright: Heating and Ventilating,
March, April, May 1940.
7.
Heat Transmission As Influenced by
Solar Radiation, by F. C. Houghten
Trans., 1932, pp. 231-284.
Heat Capacity and
and others; ASHVE
8. Heat Gain Through Walls and‘RooEs as Affected by Solar
Radiation, by F. C. Houghten and others; ASHVE Trans.,
1942, Vol. 48, pp. 21-105.
9.
Heat Loss Through Basement Walls and Floors, by F. C.
Solar Heat Gain Through Walls and Roofs for Cooling
Load Calculations, by James P. Stewart; Heating, Piping
and Air Conditioning, August 1948.
7.
8.
9.
10.
Comparative Resistance to Vapor Transmission of Various
Building Materials, by L. V. Teesdale; ASHVE Trans.,
Vol. 49, 1943, p. 124.
The Relation of Wall Construction to Moisture Accumulation in Fill Type Insulation, by Henry J. Barre; Research
Bulletin 271, Agricultural Experiment Station, Iowa State
College: Ames, Iowa, April 1940.
Vapor Transmission Analysis of Structural Insulating
Board, by F. P. Rowley and C. E. Lund: Bulletin No. 22,,
University of Minnesota Engineering Experiment Station,
October 1944.
The Diffusion of Water Vapor Through Various Building
Materials by J. D. Babbitt; Canadian Journal of Research,
Vol. 17, Sec. A, pp. 15-32, February 1939. Also Permeability of Building Paper to Water Vapor, by J. D.
Babbitt; Canadian Journal of Research, Vol. 18, Sect. A,
May 1940, pp. 90-97.
Comparison of Methods for the Determination of Water
Vapor Permeability, by Sears, Schlagenhauf, Givens and
Yett; Paper Trade Journal, Vol. 118, TAPPI Section, pp.
27-28, January 20, 1944.
Comparison of Methods for the Determination of Water
Vapor Permeability, by C. J. Weber; Paper Trade Journal,
Vol. 118, TAPPI Section, pp. 24-26, January 20, 1944.
International Critical Tables.
Vapor Barriers with Annotated Bibliography, by J. Louis
York, University of Michigan, for Office of Production,
Research and Development, War Production Board,
Washington, D. C., February 1, 1945.
How to Overcome Condensation in Building Walls and
Attics, by L. V. Teesdale; Heating and Ventilating, April
1939.
Moisture Condensation in Building Walls, by H. W.
Wooley;
National Bureau of Standards, Report BMS63.
1940.
CHART 2
1.
Preventing Condensation on Interior Building Surfaces,
by P. D. Close: ASHVE Trans., Vol. 36. 1930.
2. Permissible Relative Humidities in Humidified Buildings,
by P. D. Close; Heating, Piping and Air Conditioning,
December 1939, p. 766.
3. Methods of Moisture Control and Their Application to
Building Construction, by F. B. Rowley, A. B. Algren, and
C. E. Lund: University of Minnesota Engineering Experiment Station, Bulletin No. 17.
4. Condensation Within Walls, by F. B. Rowley, A. B. Algren, and C. E. Lund; ASHVE Trans., Vol. 44, 1938.
l-154
I’:\K’I’
CHAPTER 6
2.
Thermal Exchanges Bctwcen the Human Body ant1 Its
.\tmospheric
Environment, hy F. C. Houghten, W. W.
Tcague, W. E. Miller and W. I’. Yant; American Journal
of Physiolog)~, Vol. 88, 1929, p. 386.
3.
Heat and Moisture Losses From Men at Work and Application to Air Conditioning Problems, by F. C. Hough.
ten, W. W. Tcague, W. E. Miller and W. 1’. Yant; ASHTIE
Tf-arts., Vol. 37, 1931, p. 54.
Air Contlitioning in Industry, by W. L. Fleischer, A. E.
Staccy, Jr., F. C. Houghten and M. B. Ferberber;
ASHVE
Trans., Vol. 45, 1939, p. 59.
Pl~psiological
Basis of Medical Practice, Best and Taylor.
TABLES 41, 43 AND 44
I.
The Infiltration Problem of Multiple Entrances, Iiy j\. M.
S i m p s o n and K . I\. A t k i n s o n ; ff-lcntitig, Pii)ing nntl A i r
L’orulilio?xifq,
.Jttne 1936. p, 345.
2 . Air Lcakagc Sttttlics on Metal Windows in a Modern
OfFice Building, l)y F. C. Houghten and M. E. O’Connell;
A.SI-IJ’l< Trnns., Vol. 34, 1928, p. 321.
of Rolled Section Steel Windows,
‘.I The ~Vcathcrtightncss
Iiy J, E . llmswilcr ant1 IV. C. R a n d a l l ; A S H V E Tra?rs.,
Vol. 34. 1928, p. 527.
4. El&t of Frame Calking and Storm Windows on Inliltmt i o n Aronntl a n d T h r o u g h W i n d o w s , b y \V. ;\I. Richtmann and C. Bmatz; ASHVE Trans., Vol. 34, 1928, p. 547.
5 . Air Inliltration T h r o u g h D o u b l e - H u n g W o o d !Vindows,
by G . L . L a r s o n , D . IV. N e l s o n a n d R . W . Kubasta;
ASHVE Trans., Vol. 37, 1931. p. 571.
6. I’rcssurc Differences :\cross Windows in Relation to Wind
Velocity, by J. E. Emswiler and W. C. Randall: ASHVE
Trans.. Vol. 36, 1930, p. 83.
7 . Air Infltration Through Steel Frame Windows, by D. 0.
Rusk, V. H. Cherry and L. Boeltcr; ASHVE Trans., Vol.
39, 1933, p. 169.
4.
5.
6.
Tnlrlcs, Factors and Formulas for Computing Respiratory
Exchange and Biological Transformations of Energy, by
Thornc
M. Carpenter: Carnegie Institution, Washington,
I). c .
TABLE 49
1. \Vestinghouse
current.
data (1952) for 110~ to 125v,
60 cycle, a.c.
.
TABLE 50
I. Exhaust Hoods, by J. M. Dalla Valle.
2.
TABLE 45
1. Code of Minimum Requirements for Comfort Air Conditioning; ASHVE Trans., Vol. 44. 1938, p. 27.
2 . Ventilation Requirements, by C. I?. Yaglou, E. C. Riley
and D. J. Coggins; ASHVE Trans., Vol. 42, 1936, p. 133.
3. Control of Physical Hazards of Anesthesia, by R. M.
Tovell and A. W. Friend; Canadian Medical Association
Journal, 46560, 1942.
4 . Air Conditioning Requirements of An Operating Room
and Recovery Ward, by F. C. Houghten and W. Leigh
Cook Jr.; ASHVE Trans., Vol. 45, 1939, p. 161.
5. Code of Minimum Requirements for Heating and Ventilating Garages; ASHVE Trans., Vol. 41, 1935, p. 30.
I. LOAD ESTIM:\‘I‘ING
Reducing Heat Loads in Industrial Air Conditioning, by
L. R. St. Onge; Refrigerating Engineering, January 1946,
p. 35.
TABLE 51
1. Helpful Hints to Fried Food Fame, by the Edison General
Electric Appliance Company.
.
TABLE 53
1.
hIotors
and Generators, National Electric Manufacturers
Associa,tion Standards Publication, No. MG-1 - 1955, Part
4, p. 10.
TABLE 54
1. Heat Loss from Copper Piping, by R. H. Heilman; Heating, PiPing and Ai Conditioning, September 1935, p. 458.
CHAPTER 7
TABLE 48
1.
Heat and hloisture
Losses From the Human Body and
Their Relation to Air Conditioning Problems, by F. C.
Houghten, W. W. Teague, \V. E. Miller and W. P. Yant;
ASHVE Trans., Vol. 35, 1929, p. 245.
CHAPTER 8
Rational l’sychrometric
Formulae, by Willis H. Carrier; ASME
Trans.. Vol. 23, 1911, p. 1005.
Psychrometric
Factors in the Air Conditioning
C. M. Ashley; ASHVE Trans., Vol. 55, 1949.
Estimate,
by
Part 2
AIR DISTRIBUTION
2-1
CHAPTER 1. AIR HANDLING APPARATUS
This chapter describes the location and layout of
air handling apparatus from the outdoor air intake
thru the fan discharge on a standard air conditioning system. Construction details arc also included
for convenience.
Air handling apparatus can be of three types:
(1) built-up app aratus where the casing for the conditioning equipment is fabricated and installed at
or near the job site; (2) fan coil equipment that is
manufactured and shipped to the job site, either
completely or partially assembled; and (3) self-cont?. sd equipment which is shipped to the job site
CL sJietely assembled.
This chapter is primarily concerned with built-up
apparatus; fan coil and self-contained equipment
are discussed in Part 6. In addition to the built-up
apparatus, items such as outdoor air louvers, dampers, and fan discharge connections are also discussed
in this chapter. These items are applied to all types
.
of apparatus.
Equipment location and equipment layout must
be carefully studied when designing air handling
apparatus. These two items are discussed in detail
in the following pages.
LOCATION
The location of the air handling apparatus directly influences the economic and sound level
asr.pcts of any system.
ECONOMIC
CONSIDERATION
The air handling apparatus should be centrally
located to obtain a minimum-first-cost system. In a
few instances, however, it may be necessary to locate
the apparatus, refrigeration machine, and cooling
.tower in one area, to achieve optimum system cost.
When the three components are grouped in one
location, the cost of extra ductwork is offset by the
reduced piping cost. In addition, when the complete
system becomes large enough to require niore than
one’ refrigeration machine, grouping the mechanical
equipment on more than one floor becomes practical. This design is often used in large buildings.
The upper floor equipment handles approximately
the top 20 to 30 floors, and the lower floor equipment is used for the lower 20 to 30 floors.
Occasionally a system is designed requiring a
grouping of several units in one location, and the
use of a single unit in a remote location. This condition should be carefully studied to obtain the op
timum coil selection-versus-piping cost for the remotely located unit. Often, the cost of extra coil
surface is more than offset by the lower pipe cost for
the smaller water quantity resulting when the extra
surface coil is used.
SOUND LEVEL CONSIDERATIONS
It is extremely important to locate the air handling apparatus in areas where reasonable sound
levels can be tolerated. Locating apparatus adjacent
to conference rooms, sleeping quarters and broadcasting studios is not recommended. The following
items point up the conditions that are usually created by improper location; these conditions can be
eliminated by careful planning when making the
initial placement of equipment:
1. The cost oE correcting a sound or vibration
problem after installation is much more than
than the original cost of preventing it.
2. It may be impossible to completely correct the
sound level, once the job is installed.
3. The owner may not, be convinced even after
the trouble has been corrected.
The following practices are recommended to help
avoid sound problems for equipment rooms located
on upper floors.
1. In new construction, locate the steel floor
framing to match equipment supports designed
for weights, reactions and speeds to be used.
This transfers the loads to the building
columns.
2. In existing buildings, use of existing floor slabs
should be avoided. Floor deflection can, at
times, magnify vibration’ in the building structure. Supplemental steel framing is often necessary to avoid this problem.
3. Equipment rooms adjacent to occupied spaces
should be acoustically treated.
4. In apartments, hotels, hospitals and similar
buildings, non-bearing partition walls should
be separated at the floor and ceilings adjoining
occupied spaces by resilient materials to avoid
transmission of noise vibration.
2-2
.
5. Hearing walls adjacent to equipment rooms
should be acoustically treated on the occupied
side of the wall.
LAYOUT
Package equipment is usually factory shipped
with all of the major equipment elements in one
unit. With this arrangement, the installation can be
completed by merely connecting the ductwork and
assembling and installing the accessories.
In a central station system, however, a complete,
workable and pleasing layout must be made of all
major components. This involves considerations
usually not present in the unitary equipment installation.
The shape and cross-sectional area of the air handling equipment are the factors that determine the
INSULATE
,----CASINO -
PART 2. AIK DISTRIBUTION
dimensions oE the layout. The dehumidifier asscmbly or the air cleaning equipment usually dictates
the overall shape and dimensions. A superior air
handling system design has a regular shape. A typical apparatus is shown in I;ig. 1. The shape shown
allows for a saving in sheet metal Cabrication time
and, therefore, is considered to be better industrial
design. Its clean lines give a more workmanlike
appearance. From a functional standpoint, an irregular shaped casing tends to cause air stratification and irregular flow patterns.
The most important rule in locating the equip
ment for the air handling apparatus is to arrange
the equipment alon,u a center line for the best
air flow conditions. This arrangement keeps plenum
pressure losses to a minimum, and is illustrated in
Fig. 1.
.
-
.
ELEVATION
-.
F1c.1 -TYPICALCENTRALSTATIONEQUIPMENT
CH/\I’TER
I .
AIR H/\NDl,ING
23
:\l’I’.\R,\TUS
EQUIPMENT
This section describes available central station
apparatus cquipmcnt
ant1 recommends suitable application of the various components.
OUTDOOR AIR LOUVERS AND SCREEN
I;ig. 2 illustrates outdoor air louvers that tninimizc
the entry of snow and water into the equipment. It
is impossible to completely eliminate all moisture
with vertical louvers, and this is usually not neccssary. The screen is added to arrest most foreign
materials such as paper, trash and birds. Often the
type of screen required and the rncsh are specified
hy codes.
The screen and louver is located sufficiently above
the roof to minimize the pickup of roof dust and
the probability of snow piling up and subsequently
entering the louver during winter operation. This
height is tlcterminctl by the annual snowfall. HOWever, a minimum of 2.5 feet is recommended for
most areas. In some locations, doors are added
outside the louver for closure during extreme in-
SCREEN AND BRACES
MATERIAL
SPECIFICATIONS
Over-all Height
Maximum Over-all Width
Blades
Frame
Screen
Screen Frame
Braces
*Equivalent strength aluminum may be
Maximum
91%”
95”
22 U.S. gage steel*
18 U.S. gage steel*
I/*” #16 wire mesh
1” x 1” x l/g” angle
1” x l/g” band iron
substituted.
LOUVER WIDTH
(in.)
0 - 30
31-47
48 - 60
61 -95
Over 95
NUMBER
O F SCREENSt
NUMBER
O F BRACES1
1
0
1
1
2
1
2
2
2 equal length louvers
j-Screens over 60” high have center horizontal stiffening
braces of 1” x 1” x l/g” angle.
JBraces spaced evenly on front and back of louver and
tack welded to blade edges.
FIG. 2 - O UTDOOR A IR LOUVER AND SCREEN
clement weather suc11 as hurricanes and blizzards.
It is best to locate the outdoor air louver in such
a manner that cross contamination from exhaust
fan to louver does not occur, specifically toilet and
kitchen exhaust. In addition, the outdoor air intake
is located to minimize pulling air over a long stretch
of roof since this increases the outdoor air load
during summer operation.
Chnrt I is used to estimate the air pressure loss
at various face velocities when the outdoor louvers
are constructed, as shown in Fig. 2.
There are occasions when outdoor air must be
drawn into the apparatus thru the roof. One convenient method of accomplishing this is shown in
Fig. 3. The gooseneck arrangement shown in this
figure is also useful for exhaust systems.
.OUVER DAMPERS
The louver damper is used for three important
functions in the air handling apparatus: (1) to
control and mix outdoor and return air; (2) to
bypass heat transfer equipment; and (3) to control
air quantities handled by the fan.
Fig. 4 shows two damper blade arrangements. The
single action damper is used in locations where
the damper is either fully open or fully closed. A
double-acting damper is used where control of air
flow is required. This arrangement is superior since
the air flow is tl~rotticd more or less in proportion
to the blade position, whereas the single action
type damper tends to divert the air and does little
or no throttling until the blades are nearly closed.
Outdoor and return air dampers are located
that good mixing of the two air streams occurs.
installations that operate 24 hours a day and
located in a mild climate, the outdoor damper
occasionally omitted.
With the fan operating and the damper fully
closed, leakage cannot be completely eliminated.
C/tart 2 is used to approximate the leakage that
occurs, based on an anticipated pressure difference
across the closed damper.
Table 1 gives recommendations for various louver
dampers according to function, application, velocities and type of action required.
PITTSBURGH
PITTSBURGH
LOCK SEAM
ON 12” CENTERS
ABOVE ROOF DEPENDS
ON ANTICIPATED
FLASHING WITH WELD
CORNER SEAMS
NOTE:
Supplemental
wind bracing may be required on larger intakes.
FIG. 3
so
On
are
is
- GOOSENECK OUTSIDE AIR INTAKE
l
CI-i,\I'TlCI<
I. AIR H;\NI~I.IN(;
2-s
:\I'I',\I<.\'I'US
CHART l--LOUVER PRESSURE DROP
CHART 2-LOUVER DAMPER LEAKAGE
‘“I
/
/
.2
/
1.0
/
.9
/
/
/
.6
/
/
Al-
.02
.Ol
200
250 300
300
400
500 600 7w
7w 800
800
FACE VELOCITY (FPM)
1000
.. IO
EXAMPLE
20
30
L E A K A G E R A T E (CFM/SO
40
FT)
Solution; Pressure loss=.067 in. rg
TABLE 1 --LOUVER DAMPERS
FUNCTION OR
LOCATION
t
I
urn Outdoor Air
APPLICATION
I
VELOCITY*
(fom)
Ventilation
500-800
The higher limit may be used with short outdoor air duct connection and long return air
duct. May be single acting damper.
Permissible system
resistance and balance
500-800
Should be double acting when used for
throttling.
500-800
Single acting damper may be used.
800-1200
May b e h i g h e r v e l o c i t y w i t h s h o r t r e t u r n
duct and long outdoor air duct. Should be
double octina damoer.
400-800
Should equal cross-sectional area of dehumidifier. Should be a double acting damper.
15DO-2500
Should balance resistance of dehumidifier plus
humidifier face damper. Should be double
actina.
1000-1500
Should balance resistance
be double acting.
I
Maximum Outdoor Air
All Outdoor Air
I
Dehumidifier
Permissible system
resistance and balance
I
Permissible system
resistance and balance
Return Air
Control space
Face
conditions
I
Dehumidifier
Heater
Bypass
System balance
Bypass
Balance
I
Fan Suction or Discharge
or located in Duct
REMARKS
I
I
Available duct area
* Recommend velocity through (I fully open damper.
same or duct
Use
double
acting
at
damper.
heater.
Should
2-6
I’:\lIT
7
“.
.\IK
DISTRIl~lJTlON
CLEARANCE=;BLADE
W I T H P L U S I$
CONNECTOR
BLADE
LINKAGE
ROD ( 2 REQ’D FOR
DAMPERS OVER
38" WIDE)
ATTACHMENT
HOLES
FOR DAMPER MOTOR1
LINKAGE OR QUADRANT
DOUBLE ACTING
PARTIALLY
OPEN
SINGLE ACTING
CLOSED
SINGLE LOUVER DAMPER
SET AT
ANGLE
45'
L 7 US. GAGE STEEL
PLATE
v=\
CONTINUOUS. REOUIRED
FOR SPANS OF 12’0R MORE
HAT
CHANNEL
SECTION B -6
SECTION A-A
MULTIPLE LOUVER DAMPER ASSEMBLY
(FOR
ASSEMBLY
EXCEEDING
MAXIMUM
DIMENSIONS)
BLADES
MATERIAL
Maximum Over-all Height
Maximum Over-all Width
Maximum Blade Width
Frame - Top and Bottom
- Sides
Blades
Bearing
Blade Linkage Rod
Trunnion
Blade Link (hfulti-section)
SPECIFICATIONS
911/?”
50”
12”
3” x l/s” flat bar
3” x 7/s” x i/8” hat channel
16 U.S. gage steel
Oil-retaining porous bronze
5/ 16” dia. CRS
Die-formed
steel
Stainless steel bar
FIG.
DAMPER HEIGHT
(in.)
NUMBER
OF BLADES
To and incl. 12.11/16
12% thru 211/
219/,, thru 311/s
31s/,, thru 41%
41o/rs thru 511/2
1
2
3
4
5
519&
Sltj/,,
719/&
81D/is
6
7
8
9
thru
thru
thru
thru
611/s
711/s
811/s
911/s
4--LOUVERDAMPERARRANGEMENTS
CHAI’TEli
I . ;\IK H A N D L I N G AI’I’.\KA’TUS
\ STOP
2-7
HAT CHANNEL
/’
/,
I
i ‘--- A L U M I N U M S P A C E R
I
WASHER
I
,-LEAF
CONNECTOR
INTERCONNECTING
- STOPS
\=
“4, DIA.
ATTACHMENT
HOLES
SINGLE RELIEF DAMPER
INTERCONNECTING
SPOT WELD TO
1IBLADE
BLAD E
CONNECTION
7 U.S. GAGE PLATE
CONTINUOUS. REQUIRED
ONLY FOR SPANS
OF 12’ OR MORE
SECTION
/ 1
ALUMINUM
%A$$
1
LHAT CHANNEL
SECTION
A-A
B-B
MULTIPLE RELIEF DAMPER ASSEMBLY
(
FOR
ASSEMBLY
EXCEEDING
MAXIMUM
DIMENSION)
PRESSURE
MATERIAL
Maximum Over-all Height
Maximum Over-all Width
Maximum Blade Width
Frame - Top and Bottom
- Sides
Blades
Blade Linkage Rod
Spacer Washer
DROP
SPECIFICATIONS
91%”
40”
31/2’1
3” wide, 11 gage black iron
3” x 7/g” x l/g” hat channel
22 B & S gage aluminum
I/~” wide, 0.050” aluminum
s/s” ID x l/s” OD aluminum
FIG .
5
FACE
- RELIEF DAMPER
VELOCITY
, (fpm)
PRESSURE
DROP
(in. wg)
400
500
600
.067
.084
,120
700
800
900
.160
.200
.256
RELIEF DAMPERS
Figz~re 5 shows a typical relief damper. This accessory is used as a check damper on exhaust systems,
and to relieve excess pressure from the building.
AIR CLEANING EQUIPMENT
A variety of air filtering devices is available, each
with its own application. The pressure drop across
these devices must be included when totaling the
static pressure against which the fan must operate.
Filters are described in detail in Part 6.
HEATING COILS
Heating coils can be used with steam or hot water.
They are used for preheating, and for tempering or
reheating. The air velocity thru the coil is determined by the air quantity and the coil size. The size
‘ray also be determined by a space limitation or by
the recommended limiting velocity of 500 to 800
fpm. The number of rows and fin spacing is determined by the required temperature rise. Manufacturer’s data lists pressure drop and capacity for easy
selection.
Steam coils must be installed so that a minimum
of 18 in. is maintained between the condensate outlet and the floor to allow for traps and condensate
piping.
Preheat Coils
Non-freeze coils are recommended for preheat
service, particularly if air below the freezing temperature is encountered. To reduce the coil first
cost, the preheater is often sized and located in only
the minimum outdoor air portion of the air handling apparatus. If a coil cannot be selected at the
required load and desired steam pressure, it is better
o make a selection that is slightly undersize than
one that is oversize. An undersized coil aids in preventing coil freeze-up.
The use of two coils for preheating also minimizes the possibility of freeze-up. The first coil is
deliberately selected to operate with full steam pressure at all times during winter operation. In this
instance, the air is heated from outdoor design to
above the freezing temperature. The second coil is
selected to heat ,the air from the freezing temperature to the desired leaving temperature. The temperature of the air leaving the second coil is automatically controlled. Refer to Pm-t 3, “Freeze-up
Protection.”
In addition to the normal steam trap required to
drain the coil return header, a steam supply trap
immediately ahead of the coil is recommended.
These traps must he locatcd outsitlc the apparatus
casing.
Most coils are manufactured with a built-in tube
pitch to the return header. If the coil is not constructed in this manner, it must be pitched toward
the return header when it is installed.
To minitnirc coil cleaning problems, filters should
be installed ahead of the preheaters.
Reheat or Tempering Coils
Coils selected for reheat service are usually ovcrsized. In addition to the required load, a liberal
safety factor of from 15% to 25% is recommended.
This allows for extra load pickup during early
morning operation, and also for duct heat loss
which can be particularly significant on long duct
runs.
These coils are similar to preheat coils in that
the tubes must be pitched toward the return header.
COOLING COILS
Cooling coils are used with chilled water, well
water or direct expansion for the purpose of precooling, cooling and dehumidifying or for aftercooling. The resulting velocity thru the cooling coil
is dictated by the air quantity, coil size, available
space, and the coil load. Manufacturer’s da‘ta gives
recommended maximum air velocities above which
water carry-over begins to occur.
SPRAYS AND ELIMINATORS
Spray assemblies are used for humidifying, dehumidifying or washing the air. One item often
overlooked when designing this equipment is the
bleeder line located on the discharge side of the
pump. In addition to draining the spray heads on
shutdown, this line controls the water concentrates
in the spray pan. See Part 5, Water Conditioning.
Eliminators are used after spray chambers to prevent entrained water from entering the duct system.
AIR BYPASS
An air bypass is used for two purposes: (1) to increase room air circulation and (2) to control leaving
air temperature.
The fixed bypass is used when increased air circulation is required in a given space. It permits return air from the room to flow thru the fan without
first passing thru a heat exchange device. This
arrangement prevents stagnation in the space and
maintains a reasonable room circulation factor.
The total airway resistance for this type system
is the sum of the total resistance thru the ductwork
and air handling apparatus. Therefore, the resist-
.
CHAI’TEK I. i\ill H/\NIlLING
ante thru the bypass is normally designed to balancc the resistance of the components bypassed. This
can be accomplished by using a balancing damper
and by varying the size of the bypass opening.
The t’ollowing
formula is suggested lor use in
sizing the bypass opening :
A =
where:
2-9
i\l’I’i\KATUS
cfm
;t1
5 8 1 -\i .0707
A = damper opening (sq Et)
cEm = maximum required air quantity thru
bypass
h = design pressure drop (in. wg) thru
bypassed equipment
Temperature control with bypassed air is accom:d with either a face and bypass damper or a
P
couLrolled bypass damper alone. However, the face
and bypass damper arrangement is recommended,
since the bypass area becomes very large, and it is
difficult to accommodate the required air flow thru
the bypass at small partial loads. Even where a controlled face and bypass damper is used, leakage approaching 5% of design air quantity passes thru
the face damper when the face damper is closed.
This 5y0 air quantity normally is, included when
the fan is selected.
See Part 6 for systems having a variable air flow
to determine fan selection and brake horsepower
requirements.
FANS
Properly designed approaches and discharges
from fans are required for rated fan performance
ir
%tion to minimizing noise generation.Figures 6
U~L / indicate several possible layouts for varying
degrees of fan performance. In addition, these figures indicate recommended location of double width
fans in plenums.
Fans in basement locations require vibration isolation based on the blade frequency. Usually cork
or rubber isolators are satisfactory for this service.
On upper floor locations, however, spring mounted
concrete bases designed to absorb the lowest natural
frequency are recommended.
The importance of controlling sound and vibration cannot be overstressed, particularly on upper
floors. The number of fans involved in one location
and the horsepower required for these fans directly
determines the quality of sound and vibration control needed.
Small direct connected fans, due to higher operating speed, are generally satislactorily isolaLec1 by
rubber or cork.
In addition, all types 0C fans must have Ilexible
connections to the discharge ductwork and, where
required, must have flexible connections to the intake ductwork. Details of a rccommendcd flexible
connection arc shown in Fig. 8.
Unitary equipment should be located near columns or over main beams to limit the floor deIlection. Rubber or cork properly loaded usually gives
the required deflection for efficient operation.
FAN MOTOR AND DRIVE
A proper motor and drive selection aids in long
life and minimum service requirements. Direct
drive fans are normally used on applications where
exact air quantities are not required, because ample
energy (steam or hot water, etc.) is available at more
than enough temperature difference to compensate
for any lack of air quantity that exists. This applies, for example, to a unit heater application.
Direct drive fans are also used on applications where I
system resistance can be accurately determined.
However, most air conditioning applications use
belt drives.
V-belts must be applied in matched sets and used
on balanced sheaves to minimize vibration problems and to assure long life. They are particularly
useful on applications where adjustments may be
required to obtain more exact air quantities. These
adjustments can be accomplished by varying the
pitch diameter on adjustable sheaves, or by changing
one or both sheaves on a fixed sheave drive.
Belt guards are required for safety on all V-belt
drives, and coupling guards are required for direct
drive equipment. Figure 9 illustrates a two-piece belt
guard.
The fan motor must be selected for the maximum
anticipated brake horsepower requirements of the
fan. The motor must be large enough to operate
within its rated horsepower capacity. Since the fan
motor runs continuously, the normal 15% overload allowed by NEMA should be reserved for drive
losses and reductions in line voltages. Normal torque
motors are used for fan duty.
APPARATUS CASING
The apparatus casing on central station equipment must be designed to avoid restrictions in air
flow. In addition, it must have adequate strength
to prevent collapse or bowing under maximum operating conditions.
,
2-10
I’.\I<‘I‘
SINGLE INLET FAN
SINGLE INLET FAN
2.
.\lli
I)Is’I’l~Il~I1
‘ION
DOUBLE INLET FAN
l-FT77c’““~
l-F-/
&=J
NOTES I+2
NOTES l+2’
PLAN VIEW
PLAN VIEW
- F
PLAN VIEW
1
/
.
i‘E NOTE 3
ELEVATION VIEW
ELEVATION VIEW
DIMENSIONS :
C= DIAMETER OF FAN INLET
0.1; X C
ELEVATION VIEW (SECTION)
E = 45O MAXIMUM 30° PREFERRED
F 8 36” MINIMUM FOR ACCESS DOOR
.
INLET CONNECTIONS ’
BEST
GOOD
T R A N S F O R M A T I O N I” I N 7 ” P R E F E R R E D ,
I” IN 4” PERMITTED NOTE 68 7
DIMENSIONS :
A.l;xB TO 23x8
6 = FAN DISCHARGE OPENING,
LARGEST DIMENSION
DISCHARGE CONNECTIONS
NOTES:
1. Fan should be centered in casing to provide good flow
conditions.
5. Use square vaned elbow for best results, with take-off
in opposite direction to fan rotation.
2. All equipment should be centered for best performance.
G. Slope of 1” in 4” recommended for low velocity.
3.
7. Slope of 1” in 7” recommended for high velocity.
Angle “E” is used to determine “F” distance between
equipment and fan.
4 . R, = G” minimum. Vane spacing determined from
Chart 6.
F IG.
(i-SINGLEFAN INLET
,
AND
DISCHARGECONNECTIONS
(:I-I.\l”I‘l~l<
I.
.\II<
II.\NI)I.IN(;
2-11
.\I’I’.\I<,\~I’IJ,S
‘E I
r NOTE I
NOTE I
NO
TYPICAL
VANE
LOCATION
R, = 6: R2OETERMINED
FROM CHART 6. VANE SPACING
A .I$B T O 24B
B = LONGEST DIMENSION OF OUTLET OPENING
NOTES:
1. Transformations to supply duct have maximum slope of
1” in 7”.
3. Do not install ducts so that the air flow is counter to
fan roration. If necessary, turn fan section.
2. Square elbows with double thickness vanes may be
substituted.
4. Transformations and units shall be adequately supported so no weight is on the flexible fan connection.
I’,\i<.I‘
2-F
2.
\ II<
I~IS’I‘l~II11J~I‘ION
I” FLANGE AND HEM
ALTERNATE
POSITION OF
B O L T ~-
IMPREGNATED FABRIC
- 20 U.S. GAGE STEEL -
RECTANGULAR (FAN OUTLET)
SEALING COMPOUND APPLIED
BETWEEN FLEXIBLE CONNECTION
AND FAN BEFORE ASSEMBLY l
5”
16 F L A N G E
BOLT ON 4” CENTERS
SEALING COMPOUND APPLIED
BETWEEN FLEXIBLE CONNECTION
AND CASING BEFORE ASSEMBLY *
SHEET METAL SCREWS
ON 12” CENTERS
r
BAND
.
--‘--x
I ” x ;’ B A N D I R O\N -
-SEALING
*REQUIRED ON HIGH
PRESSURE SYSTEMS
COMPOUND APPLIED BETWEEN
FABRIC, BAND AND 18 US. GAGE STEEL,
BEFORE ASSEMBLY * \
ONLY,
ROUND
(F
A N
INLET)
Frc.8 - FLEXIBLE CONNECTIUN
SHEAVE
SECTION A-A
A.B.C.R,.Rp,+ R3 DIMENSIONS
REQUIRED FOR CONSTRUCTION
EXPANDED METAL AND
1” HEM REVERSED IN
THIS SECTION
S
T
A/ R
L CENTER LINE
OF SHEAVE AT
U
P
T
I
E;ich sheet ol m:ltcri;ll sl10uItl he l’:~l)ricatctl a s a
panel and joined together, ;ks illrlstratetl i n Pig /0,
by standing sc~iiiis, bolted or riveted on 12 in. kentars. Nornially, SC’:IIIISperpendicular to air flow arc
i>l;lcecl outsitlc of the casing. Side walls over 6 It
high ;intl roof spns over fi I t witlc require supplemental reinforcing as shown in Tol~le -3. Diagonal
angle braces as illustratcrl
in Pig. 1 I may also lx
rcquiretl.
The rccommcntled construction of apparatus
casings and connections between equipment components (except when mounted in the ducts) is 18
U.S. gage steel or 16 13 PCS gage aluminum. Aluminuni
in contact with galvanized steel at connections
to spray type equipment requires that the inside of
the casing be coat& with an isolating material for a
distance of 6 in. from the point of contact.
CONNECTIONS TO MASONRY
I\ coricretc curl) is rccomrncntletl t o p r o t e c t insul;itiori I’rorii clcterior;kting whcrc t h e aplxir:ltits
casing joins the Iloor. I t a l s o provitlcs ;I iI nilorm burface f o r ;itt:iching t h e casing; t h i s conserves 1’:ibriT
cation time. Figure I- illustrxtes the rcconimcntlctl
methotl OF attaching a casing to the curl).
When an equipment room wall is used as one sitlc
of the ;ipparxtus, the casin,cr is attachctl 2s shown in
Fig. 13.
The degree oE tightness required for an appratus
casing tlcpentls on the air conditioning application.
DE
FIG. I1 - APPARATUS CASING
FIG. 10 - APPARATUS CASING S EAMS
TABLE
2--SUPPLEMENTAL
REINFORCING
REINFORCING
FOR
APPARATUS
SIDE WALL HEIGHT
OR ROOF WIDTH
w
NUMBER OF
ANGLES*
ANGLE SPACING
6t0 a
Et0 12
1
2
middle
‘Y3 points
-
over 12
variable
4 ft centers
3 & 4 panels
5 & 6 panels
7 a 0 panels
‘For lengths up to 12 ft., use 1 % x 1 ‘/z x ‘/a in. angle.
CASTING LENGTHS
For lengths over 12 ft., use I 3/4 x 13/ x %6 in. angle.
CASING
DIAGONAL
ANGLE BRACES*
(pairs)
1
2
3
2-14
l’;\l<-I‘
2 .
:\IK I~IS~I’RIlIIJ-I‘ION
\
For instance, 011 a pull-thru system, leakage between
the tlehuniiclificr ant1 the ran cannot be tolcratetl if
the aplxirattls is in a humid non-contlitionetl space.
.-11~0, as the negative pressure at the fan intake in,crcascs, the less the leakage that can I)e tolcratctl.
If the apparatus is located in a return air plenum,
normal construction as shown in I;ig.~. 1-3 ~ntl /3
cm be used. Corresponding construction practice
for equipment requiring extreme care is shown in
18 U.S. GAGE STEEL CASING
Ifx
Figs. 14, 15 rind Ih.
EXPANSION
BOLT
ON 12” CENTERS
FIG.
12 - CONNECTION
TO
M ASONRY CURB
In addition to the construction required for lcakage at seams, pipes passing thru the casing at cooling coil connections must bc scaled as sl~ow~i in
Fig. f7. This applies in applications where the temperature difference between the room and supply
Gr temperatures is 20 F ant1 greater.
DRAINS AND MARINE LIGHTS
Upkeep and maintenance is better on a.11 appxthat can be illuminated and easily cleaned than
on one that does not have good illumination ant1
drainage. To facilitate this maintenance, marine
lights, as well as drains, arc recommended as shown
in Fig. 1.
As a rule of thumb, drains should be located in
the air handling apparatus wherever water is likely
to accumulate, either in normal operation of the
equipment or because of maintenance. Specific
exampIeS are:
1. In the chamber immediately after the outdoor
air louver where a driving rain or snow may
accumulate.
2. Before and after filters that must be periodically washed.
3. Before and after heating and cooling coils that
must be periodically cleaned.
4. Before and after eliminators because of backlash and carry-over due to unusual air eddies.
Drains should not normally be connected directly
to sewers. Instead, an open site drain should be used
as illustrated in Pn~t 3.
tus
CASING RIVETED
ON APPROXIMATELY
12” CENTERS
ANGLE INSIDE
OF CASING,
‘EXPANSION
FIG.
13
-
BOLT ON 12” CENTERS’
CONNECTION
TO
M ASONRY W ALL
APPLY SEALING COMPOUND
TO ANGLE BEFORE
RIVET ON 6” CENTERS
EXPANSlON
FIG. 1-I - Low
DEWPOINT
BOLT ON 12” CENTERS
MASONRY
CONNECTIONS
.
CURB
INSULATION
Insulation is required ahead of the preheater ant1
vapor sealed for condensatidn during winter operation. Normally, the section of the casing from the
preheater to the dehumidifier is not insulated. The
dehumidifier, the fan and connecting casing must
be insulated and vapor sealed; fan access doors are
not insulated, however. The bottoms ant1 sides of
the dehumidifier condensate pan must also be insulated, and all parts of the building surfaces that
I
(:11,\1”1’1-1< 1. /\IK lI,\KI)I.IN(~
2-15
,\1’l’.\l</\‘I‘IJS
SERVICE
J$luil)iiient service is csscntial :lntl sp;lcC mllst bc
provided to acco~nplish
this service. It is recommcntlctl th;tt minimuln cle:lr;inccs
be m a i n t a i n e d
so tll;tt ;ICCCSS to dI equipment is ;lvailablc. In x&Iitioli, provision sl~oultl be made so that equipment
can be removed without dismantling the complete
apparatus. Access must be provitlecl for heating and
cooling coils, steam tr:lps, damper motors and linkages, control valves, bearings, fan motors, lans and
similar components.
Scrvicc access doors as illustrated in Fig. 18 are
rcconimendetl, ant1 are 1oc:itcd in casing sections as
shown in Fig. 1.
conserve rioor space, the entrance to the cquipment room is olten located so that coils can be removed directly thru the equipment room doors.
This arrangement requires less space than otherwise possible.
IL‘ the equipment room is not arranged as described, space must be allowed to clean the coil
tubes mechanically. This ap’plies
to installations
that have removable water headers.
-,
APPROXlMATELY
EXPANSION
I ;$i-aj y
BOLT ON 12” CENTERS
FIG . 15 -Low DEWPOINT MASONRY WALL
CONNECTIONS
CENTERS
OUNO BRUSHED
EMBLY
FIG. I6 -
SK,\LJN(;
C
S
A
ACCESS PANELS WITH
SHEET METAL SCREWS,
SEAL WITH SEALING
COMPOUND.
APPLY SEALING
COMPOUND TO ANGLE
BEFORE R’“ET’NG
RIVET OR BOLT ON
APPROXIMATELY 12”
STANI)IN(; Skh\Ms
SEE
DETA,,.
“8”
-WINDOW
TYPE SASH.
HANDLE
BUTT
DR E A C H - T H
OR U A C C E SOS
SINGLE
R
ACTING
HANDLE
I SEE OETAIL “B”l
“A”,
SASH
LOCK
DOOR
FRAME
EXTENSION COLLAR
U S E D W H E N D U C T IS
INSULATED
D E T A I L “8”
~WINDOW SASH LOCK
EXTENSION COLLQR LENGTH (SEE DETAIL 8, IS OETERMlNED
BY INSULInON
THICKNESS USED ON 0”CT OR CASING.
.
DOOR,
/GASKET
MATERIAL SPECIFICATION
1. Door - 24 U.S. gage steel or 22 B &S gage aluminum.
CASING
2. Frame - 24 U.S. gage steel or 22 B &S gage aluminum.
3. Extension Collar - Same gage as duct metal.
4. Formed Protection Angle - 18 U.S. gage steel or 16
B & S gage aluminum.
DETAIL “A”
DOUBLE
ACTING
DOOR
HANDLES
3. Angle Brace - ls/4” x ls/a” x i/s” angle.
6. Butt Hinges - Steel.
7. Gaskets - Felt.
F O R M E D PROTECTlON
8. Fastener a. Walk-thru door: Three double acting handles.
b.
ACCESSDOOR
ANGLE
BRACE
Reach-thru
sash lock.
Single
acting
Walk-thru Access Doors
Normal size - 22” W x 58” H
Reach-thru
Access
Norma! sizes
CASING INSIDE
SECTION
door:
A-A
F IG. 18 - ACCESS
DOORS
Doors
W
H
10” x 12”
12” x 16”
16” x 04”
b
handle
with
window
2-17
CHAPTER 2. AIR DUCT DESIGN
The function of a duct system is to transmit air
from the air handling apparatus to the space to
be conditioned.
To fulfill this function in a practical manner,
the system must be designed within the prescribed
limits of available space, friction loss, velocity, sound
level, heat and leakage losses and gains.
This chapter discusses these practical design
criteria and also considers economic balance be/. . -n first cost and operating cost. In addition, it
0. .s recommended construction for various types
of duct systems.
GENERAL SYSTEM DESIGN
CLASSIFICATION
Supply and return duct systems are classified with
respect to the velocity and pressure of the air within
the duct.
Velocity
There are two types of air transmission systems
used for air conditioning applications. They are
called conventional or low velocity and high velocity systems. The dividing line between these systems
is rather nebulous but, for the purpose of this
chapter, the following initial supply air velocities
X
are offered as a guide:
1. Commercial comfort air conditioning
a. Low velocity - up to 2500 fpm. Normally
between 1200 and 2200 fpm.
b. High velocity - above 2500 fpm.
2. Factory comfort air conditioning
a. Low velocity - up to 2500 fpm. Normally
between 2200 and 2500 fpm.
b. High velocity - 2500 to 5000 fpm.
Normally, return air systems for both low and
high velocity supply air systems are designed as low
velocity systems. The velocity range for commercial
and factory comfort application is as follows:
1. Commercial comfort air conditioning - low
velocity up to 2000 fpm. Normally between
1500 and 1800 fpm.
2. Factory comfort air conditioning - low velocity up to 2500 fpm. Normally between 1800 and
2200 fpm.
Pressure
Air distribution systems are divided into three
pressure categories; low, medium and high. These
divisions have the same pressure ranges as Class I,
II and III fans as indicated:
1. Low pressure - up to 3% in. wg - Class I fan
2. Medium pressure - 3s/4 to 63, in. wg - Class
II fan
3. High pressure - ($4 to 12% in. wg - Class III
fan
These pressure ranges are total pressure, including
the losses thru the air handling apparatus, ductwork
and the air terminal in the space.
AVAILABLE SPACE AND ARCHITECTURAL APPEARANCE
The space allotted for the supply and return air
conditioning ducts, and the appearance of these
ducts, often dictates system layout and, in some
instances, the type of system. In hotels and office
buildings where space is at a premium, a high velocity system with induction units using small round
ducts is often the most practical.
Some applications require the ductwork to be
exposed and attached to the ceiling, such as in an
existing department store or existing office building.
For this type of application, streamline rectangular
ductwork is ideal. Streamline ductwork is constructed to give the appearance of a beam on the
ceiling. It has a smooth exterior surface with the
duct joints fabricated inside the* duct. This ductwork is laid out with a minimum number of reductions in size to maintain the beam appearance.
Duct appearance and space allocation in factory
air conditioning is usually of secondary importance.
A conventional system using rectangular ductwork
is probably the most economical design for this
application.
ECONOMIC FACTORS INFLUENCING DUCT LAYOUT
The balance between first cost and operating
cost must be considered in conjunction with the
available space for the ductwork to determine the
best air distribution system. Each application is
different and must be analyzed separately; only
general rules or principles can be given to guide
the engineer in selecting the proper system. The
PART 2. AIR DISTRIBUTION
2-18
CHART J-DUCT HEAT GAIN VS ASPECT RATIO
I
2. Ducts carrying small air quantities at a low
velocity have the greatest heat gain.
3. The addition of insulation to the duct dccreases duct heat gain; for example, insulating _
the duct with a material that has a U value of
.12 decreases heat gain 907”.
It is, therefore, good practice to design the duct
system for low aspect ratios and higher velocities to
minimize heat gain to the duct. If the duct is to run
thru an unconditioned area, it should be insulated.
Aspect Ratio
,Jlowing items directly inHuencc the first and operating cost:
1. Heat gain or loss from the duct
2. Aspect ratio of the duct
3. Duct friction rate
4. Type of fittings
Heat Gain or Loss
The heat gains or losses in the supply and return
duct system can be considerable. This occurs not
only if the duct passes thru an unconditioned space
but also on long duct runs within the conditioned
space. The transfer of heat takes place from the
space to the air in the duct when cooling, and from
the air to the space when heating.
An allowance must be made for duct heat gain
for that portion of the duct in the unconditioned
‘-ace when estimating the air conditioning load.
e method of making this allowance is presented
in Part I, Load Estimating. This allowance for duct
heat gain increases the cooling capacity of the air.
This increase then requires a larger air quantity
or lower supply air temperature or both.
TO compensate for the cooling or heating effect
of the duct surface, a redistribution of the air to
the supply outlets is sometimes required in the
initial design of the duct system.
The following general guides are offered to help
the engineer understand the various factors influencing duct design:
1. Larger duct aspect ratios have more heat gain
than ducts with small aspect ratios, with each
carrying the s a m e air quantity. Chart 3 illustrates this relationship.
.
The aspect ratio is the ratio of the long side to
the short side of a duct. This ratio is an important
factor to be considered in the initial design. Increasing the aspect ratio increases both the installed cost
and the operating cost of the system.
The installed or first cost of the ductwork depends
on the amount of material used and the difficulty
experienced in fabricating the ducts. Table 6 reflects these factors. This table also contains duct
class, cross-section area for various round duct sizes
and the equivalent diameter of round duct for rectangular ducts. The large numbers in the table are
the duct class.
The duct construction class varies from F to 6
and depends on the maximum duct side and the
semi-perimeter of the ductwork. This is illustrated
as follows:
DUCT
CLASS
1
2
3
4
5
6
MAX. SIDE
SEMI-PERIMETER
(in.)
6 - 171/
12-24
26-40
24 - 88
48 - 90
go-144
(in.)
10
24
32
48
96
96
-
23
46
46
94
176
238
Duct class is a numerical representation of relative first costs of the ductwork. The larger the duct
class, the more expensive the duct. If the duct class
is increased but the duct area and capacity remain
the same, the following items may be increased:
I. Semi-perimeter and duct surface
2. Weight of material
3. Gage of metal
4. Amount of insulation required
Therefore, for best economics the duct system
should be designed for the lowest duct class at the
smallest aspect ratio possible. Example I illustrates
the effect on first cost of varying the aspect ratio
for a specified air quantity and static pressure requirement.
.
CHAPTEK
2.
AIR
DUCT
2-19
DESIGN
Example 1 -Effect of Aspect Ratio on First Cost of the
Ductwork
CHART 4-INSTALLEDCOST VS ASPECT RATIO
Given:
Duct cross-section area - 5 . 8 6 xl ft
Space available - unlimited
Low velocity duct system
Find:
The duct dimensions, class, surface area, weight and gage of
metal required.
Solution:
I. Enter Tnble 6 at 5.86 sq ft and determine the rectangular
duct dimensions and duct class (see tabulation).
2. Determine recommended metal gages from TaOles 14 and
15 (see tabulation).
3. Determine weight of metal from T&/e 18 (see tabulation).
DIMENSION
ARE:\
(in.)
69 ft)
94x 12
84x 13
7Gx 14
42 x 22
30 x 30
32.5 (round)
DIMENSION
(in.)
5.86
5.86
5.86
5236
5.86
5.86
ASPECT
RATIO
7.8: 1
6.5:l
5.4: 1
1.9:1
1:l
-
SURFACE
GAGE
AREA
(U.S.) .
(sq ft / ft)
DUCT
CONSTR.
CLASS
6
5
4
4
4
WEIGHT
(lb/f9
When the aspect ratio increases from 1: 1 to 8: 1,
the surface area and insulation requirements increase 70y0 and the weight of metal increases over
three and one-half times. This example also points
oi*- ‘hat it is possible to construct Class 4 duct, for
tl . given area, with three different sheet metal
gages. Therefore, for lowest first cost, ductwork
should be designed for the lowest class, smallest
aspect ratio and for the lightest gage metal recommended.
Chart 4 illustrates the percent increase in in.stalled cost for changing the aspect ratio of rectangular duct. The installed cost of round duct is also
included in this chart. The curve is based on installed cost of 100 ft of round and rectangular duct
with various aspect ratios of equal air handling
capacities. The installed cost of rectangular duct
with an aspect ratio of 1: 1 is used as the 100% cost.
Friction Rate
The operating costs of an air distribution system
can be adversely influenced when the rectangular
duct sizes are not determinccl from the table of circular equivalents (Table 6). This table is used to
obtain rectangular duct sizes that have the same
friction rate and capacity as the cquivalcnt round
duct. For example, assume that the required duct
area for a system is 480 sq in. and the rectangular
duct dimensions arc determined directly from this
area. The following tabulation shows the resulting
equivalent duct diameters and friction rate when
4000 cfm of air is handled in the selected ducts:
DUCT
DIM.
(in.)
24 x 20
30x 16
48x 10
80x 6
1
EQUIV / F R I C T I O N j
.-\SI’ECT
R&ND
RATE
DUCT DIAM
R.+TIO
(in.)
(in. wg/lOO ft)
23.9
23.7
22.3
20.1
,090
,095
,125
,210
1.2:1
1.9:1
4.8: 1
13.3:1
If a total static pressure requirement 0C 1 in.,
based on 100 ft of duct and other equipment is
assumed for the above system, the operating cost
increases as the aspect ratio increases. This is shown
in C/MU-~ 5.
Therefore, the iowcst owning and operating cost
occurs where round or Spired-Pipe is used. II' round
2-%.I
\
CHART 5-OPERATING COST VS ASPECT RATIO
DUCT
LAYOUT
PART
2. AIR DlSTRlBUTlON
CONSIDERATIONS
There are several items in duct layout that should
he considered before sizing the ductwork. These include duct transformations, elbows, fittings, take-offs,
duct condensation and air control.
Transformations
ductwork cannot be used because of space limitations, ductwork as square as possible should be used.
An aspect ratio of 1: 1 is preferred.
Type of Fittings
In general, fittings can be divided into Class A
and Class B as shown in Table 3. For the lowest
first cost it is desirable to use those fittings shown
as Class A since fabrication time for a Class B fitting
is approximately 2.5 times that of a Class A fitting.
Duct transformations are used to change the
shape of a duct or to increase or decrease the duct
area. When the shape of a rectangular duct is
changed but the cross-sectional area remains the
same, a slope of 1 in. in 7 in. is recommended for
the sides of the transformation piece as shown in
Fig. 19. If this slope cannot be maintained, a maximum slope of 1 in. in 4 in. should not be exceeded.
Often ducts must be reduced in size to avoid obstructions. It is good practice not to reduce the duct
more than 20y0 of the original area. The recommended slope of the transformation is 1 in. in 7 in.
when reducing the duct area. Where it is impossible
to maintain this slope, it may be increased to a
maximum of 1 in. in 4 in. When the duct area is increased, the slope of the transformation is not to
exceed 1 in. in 7 in. Fig. 20 illustrates a rectangular
duct transformation to avoid an obstruction, and
Fig. 21 shows a round-to-rectangular transformation
.
to avoid an obstruction.
TABLE 3-DUCT FITTING CLASSES
CLASS
A-NO
VANED
FITTINGS
Any fitting with constant cross-section
dimensions.
FIG .
19
- DUCT TRANSFORMATION
Any fitting with changing radius and
constant width.
Fittings with straight sides and seams.
63
CLASS B-ALL
VANED
FITTINGS
Any fitting with concentric radii, and
changing width.
Any fitting with eccentric radii and
changing width.
NOTE: 1:7 slope is recommended for high velocity,
1:1 slope for low velocity.
FK.
20 - RECTANGULAR D~JCT TRANSFORMA-VION
-IX) AVOII)
OKSTRIICTION
l
2-21
CHAPTER 2. AIR DUCT DESIGN
In some air distribution systems, equipment such
as heating coils is installctl in the ductwork, Normally the equipment is larger than the ductwork
and the duct area must be incrcascd. The slope 01
the tr~~nsformation piece on the upstream sitlc ol
the equipment is limited to 30” as shown in I;ig. 22.
On the leaving sick the slope should be not more
than ,15”.
Duct Reduction Increments
L
,\cceptcti methods of duct design visually indicate
;I reduction in duct area after each terminal and
branch take-off. Unless a reduction of at least 2 in.
can be made, however, it is recommended that the
original duct size be maintained. Savings in installed cost of as much as 25% can he realized by
running the duct at the same size for several
terminals.
TOP VIEW
/I
:
I
NOTE: Angles shown
are for low velocities. I:7 slope
is recommended for high velocities.
FK. 22 - DUCT TRANSFORMA.IXON
E QUIPMENT
Locating pipes, electrical conduit, structural
members and other items inside the ductwork should
always be avoided, especially in elbows and tees.
Obstruction of any kind must be avoided inside
II
I
\
-Sk&&
L
MAX
ELEVATION
Obstructions
Ii
:
I
’ duct sizes should be even dimensions and all
reci;tctions should be in 2 in. increments, preferably
in one dimension only. The recommended minimum duct size is 8 in. x 10 in. for fabricated shop
ductwork.
1.
MAX
IN THE
WITH
DIJCI
high velocity ducts. Obstructions cause unnecessary
pressure loss and, in a high velocity system, may
also be a source of noise in the air stream.
I/
1 ‘I
1 jl
’
’ I;
IY
I
-
1
;
I
~MAXIMUM DUCT AREA
REDUCTION - 20 %
NOTE: I:7 slope is recommended for high velocity, I:4 slope for low velocity.
-
/
, PART
2-22
2. AIR DISTRIBUTION
111 those few instances in which obstructions must
1x1~s thru the duct, use the Following rccommendations:
1. Cover all pipes and circular obstructions over
‘1 in. in diameter with an easement. Two
typical easements are illustrated in Fig. 23.
2. Cover any flat or irregular shapes having a
width exceeding 3 in. with an easement.
Hangers or stays in the duct should be parallel
to the air flow. If this is not possible, they
should be covered with an easement. I;ig. 24
shows a tear drop-shaped easement covering
an angle. Hanger “B” also requires an easement.
3. If the easement exceeds 20% of the duct area,
the duct is transformed or split into two ducts.
When the duct is split or transformed, the
original area should be maintained. Fig. 25
illustrates a duct transformed and a duct split
to accommodate the easement. In the second
case, the split duct acts as the easement. When
the duct is split or transformed, slope recommendations for transformations should be
followed.
4. If an obstruction restricts only the corner of
the duct, that part of the duct is transformed
to avoid the obstruction. The reduction in duct
area-must not exceed 20y0 of the original area.
Elbows
A variety of elbows is available for round and
rectangular duct systems. The following list gives
the more common elbows:
FIG. 24 - EASEMENTS COVERING IRREGULAR SHAPES
Rectangular Duct
1. Full radius elbow
2. Short radius vaned elbow
3. Vaned square elbow
Round Duct
1. Smooth elbow
2. S-piece elbow
3. 5-piece elbow
NOTE: 1:7 slope is recommended for high velocity,
I:4 slope for low velocity.
FIG. 25
- DUCT T RANSFORMED FOR EASEMENW
FIG.
26
-
FULL RADIUS RECTANGULAR ELBOW
l
CHAP-I-EK
2 .
2-23
AIK DUC’I‘ DESIGN
The elbows are listed in order of minimum cost.
‘l’his sequence does not necessarily indicate the
minimum pressure drop thru the elbow. Tal~lk 3
thtx I2 show the losses for the various rectangular
and round elbows.
Full radius elbows (Fig. 26) are constructed with :I
throat radius equal to s/4 of the duct dimension in
the direction of the turn. An elbow having this
throat radius has an K/D ratio of 1.25. This is considered to be an optimum ratio.
The short radius vaned
r
elbow and their location is determined from Chad
6. Exnmple 2 illustrates the use of Chart 6 in deter-
mining the location of the vanes in the elbow in
Fin 28.
HEEL RADIUS
D = 20”
elbow is shown in Fig. 27.
This elbow can have one, two or three turning
vanes. The vanes extend the full curvature of the
CENTER CINE RADIUS
L
I
THROATRADIUS=3”
FIG. 28 -RECTANCULARELBOWVANELOCAIION
Exu,np/e 2 L Locating Vanes in a Rectangular E/bow
Example 3 - locating Vanes in a Rectangular Elbow with
a Square Throat
Given:
Rectanglllar elbow shown in Fig. 28.
Throat radius (I?,) - 3 in.
Duct dimension in direction of turn - 20 in.
Heel radius (2X),) - 23 in.
Find:
1. Spacing for two vanes.
2. R/D ratio of ell,ow.
Soluticn:
Given:
Elbow shown in Fig. 29.
Throat radius - none
Heel radius - 20 in.
Duct dimension in direction of turn - 20 in.
Find:
Vane
1. Enter Ckrt 6 at X, = 3 in. and I?,, = 23 in. Read vane
spacing for R, and R, (dotted line on chart).
R, = 6 in. RL = 12 in.
2. The centerline radius R of the elbow equals 13 in.
Therefore R/D = 13/20 = .65.
Although a throat radius is recommended, there
may be instances in which a square corner is imperative. Chart 6 can still be used to locate the vanes. A
th- It radius is assumed to equal one-tenth of the
ht .I adius. Example 3 illustrates vane location in
an elbow with a square throat.
r
spacing
Solution:
.\ssume a throat radius equal to .l of the heel radius:
.l x 20 = 2 in.
Enter Cl~art 6 at R, = 2, and R, = 20 in. Read vane spacing
for R, and R2.
R, = 4.5 in. R, = 9.5 in.
In addition a third vane is located at 2 in. which is the
assumed throat radius.
EEL
ADIUS (I+,)
D
I_
Rh
T H R O A T R A D I U S t R+)
F1c.27 -SHORTRADIUSVANEDELBOW
F1c.29 -RECTANGULARELBOWWITH No
THROATRADIUS
CHART 6-VANE LOCATION FOR RECTANGULAR ELBOWS
2
3
4
5
6
THROAT RADIUS (Rt) (IN.)
7 8910
20
30
40
ej0
Y
60 70 t30 90 lO(
.UU
hPLE 2
NO. I OF 3
I
i
I
\
++
btnttttt
+-I-
l++H+n
NO. I OF 2
.,
,
NO. 2 OF 3
/
-:
NO. I OF I
++
tttttfttt
*-tnttm
++
+*tttt
+t
+tUt+Hi
n7
Tnrrrm
.
.
I
2
3
4
5
6
7 8 910
HEEL RADIUS (R”) (IN3
20
’
30
40
50
60 70 609OlW
From Fan Engineering, Buffalo Forge Co.
A vaned square elbow has either double or single
thickness closely spaced vanes. Fig. 30 illustrates
double thickness vanes in a square elbow. These
lbows are used where space limitation prevents
the use of curved elbows and where square corner
elbows are required. The vaned square elbow is
expensive to construct and usuaily has a higher
pressure drop than the vanecl short radius elbow
and the standard elbow (R/D = 1.25).
Smooth elbows are recommended for round or
.$pil-n-Pipe systems. Fig. 31 illustrates a 90” smooth
elbow with a R/D ratio of 1.5. This R/D ratio
is standard for all elbows used with round or SpzmPipe duct.
FIG.
31-90” SMOOTH
ELBOW
CHAI’TEK
2 .
AIR DUC’I
2-25
DESIGN
,4 S-piece elbow (I;;g. 32) has the same R/D ratio
as a smooth elbow but has the highest pressure drop
oE either the smooth or 5-piece elbow (Fig. 33).
This elbow is second
in construction costs and
should be used when smooth elbows are unavailable.
5-piece elbow (I;i,.(7 33) has the highest first
cost of all three types, It is used only when it is
necessary to reduce the pressure drop below that of
the 3-piece elbow, and when smooth elbows are not
available.
A 45” elbow is either smooth (Fig. 34) or 3-piece
(Fig. 35). A smooth 45” elbow is lower in first cost
and pressure drop than the S-piece 45” elbow. A
3-piece elbow is used when smooth elbows are not
available.
Take-offs
There are several types of take-offs commonly
used in rectangular duct systems. The recommendations given for rectangular elbows apply to takeoffs. Fig. 36 illustrates the more common take-offs.
Fig. 36A is a take-off using a full radius elbow. In
Figs. 36A and 36B the heel and throat radii originate from two different points since D is larger than
D,.
e principal difference in Figs. 36A and 36B
is thar the take-off extends into the duct in Fig. 36B
and there is no reduction in the main duct.
Figure 36C illustrates a tap-in take-off with no
part of the take-off extending into the duct. This
FIG .
33
- 90" !&PIECE
ELBOW
FIG. 34 -,15”
SMOOTH
E~uow
type is often used when the quantity of air to be
taken into the branch is small. The square elbow
take-off (Fig. 36D) is the least desirable from a cost
and pressure drop standpoint. It is limited in application to the condition in which space limitation
prevents the use of a full radius elbow take-off.
A straight take-off .(l;ig. 37) is seldom used for
duct branches. Its use is quite common, however,
when a branch has only one outlet. In this instance
it is called a collar. A splitter damper can be added
for better control of the air to the take-off.
There are two varieties of take-offs for round and
Spira-Pipe duct systems: the 90” tee (Fig. 38) and the
90” conical tee (Fig. 39). A 90” conical tee is used
when the air velocity in the branch exceeds 4000
fpm or when a smaller p r e s s u r e d r o p t h a n t h e
straight take-off is required. Crosses with the takeoffs located at ISO”, 135” and 90” to each other are
shown in Fig. 40.
When the duct system is designed, it may be necessary to reduce the duct size at certain take-offs. The
reduction may be accomplished at the take-off (Fig.
41) or immediately after the take-off (Fig. 42). Reduction at the take-off is recommended since it eliminates one fitting.
FIG .
35 - 45’ ~-PIECE E LBOW
l'.\l<'l‘ 2. ,\Il< I)IS'I'KII~UTION
2-s
i
FIG. 36 - TYPICALTAKE-OFFS
I,
-7
AIR
FLOW
c
I,
FIG. 37 -OUTLETCOLLAR
r
FIG. 38 - 90” TEE
(;FI~\1’T1:II
2.
\II<
I>IJ(:‘I‘
I)I~:SI(;N
2-27
TABLE 4-MAXIMUM DIFFERENCE BETWEEN SUPPLY AIR TEMPERATURE AND ROOM DEWPOINT
WITHOUT CONDENSING MOISTURE ON DUCTS (F)
AIR VELOCITY IN STRAIGHT RUN OF DUCT (FPM)*
AIR CONDlTlONS
-‘tRROUNDING
DUCT
Bright
Bright
Bright
P a i n t e d M e t a l P a i n t e d M e t a l P a i n t e d Metal
400
74 - 100
I
55
60
70
80
85
VALUE OF $ _ ,
15
15
18
15
13
13
11
10
11
10
7
4
3
7
4
3
13
9
6
4
1
800
20
.90
1 -
.66
1 .66
1200
8
7,
5
3
2
6
A
3
2
.A2
1
.49
2000
1600
8
7
4
5
4
2
2
11
10
9
;
1
Bright
Bright
Bright
Painted Metal Painted Metal Painted Metal
1
8
7
6
5
4
2
2
.31
1
5
5
A
3
2
2
7
6
5
A
3
2
2
1
.37
1 .2A
1
.31
3000
A
A
3
3
2
5
A
A
3
2
2
1
1
1 .20
3
3
2
2
2
1
1
1
1 .23
1 .15
*For elbows and other fittings, see Notes 4 and 7.
E Q U A T I O N : tdp - ts, = (trm - tdp) (;where:
1 )
tdp
= duct surface temp. assumed equal to room dewpoint.
= supply air dry-bulb temp in duct.
tsa
trm = room d r y - b u l b t e m p .
NOTES:
1. Exceptional Cases: Condensation will occur at a lower relative
humidity than indicated in the table when f;! falls below the average value of 1.65 for painted ducts and 1.05 for bright metal
ducts. The radiation component of fz will decrease when the
ct is exposed to surfaces colder than the room air, as new a
d wail. The convection component will decrease for the top
of ducts, and also where the air flow is obstructed, such as a
duct running very close to a partition. If either condition exists,
use value given for CI relative humidity 5% less than the relative humidity in the room. If both conditions exist, use value
given for 10% lower relative humidity.
2. Source: Calculated using film heat transmission coefficient on inside of duct ranging from 1.5 to 7.2 Btu/(hr) (sq ft) (deg F).
The above equation is based on the principle that the temperature drop through any layer is directly proportional to its
thermal resistance. It is assumed that the air movement surrounding the outride of the duct does not exceed approximately 50
fpm.
3. For Room Conditions Not Given: Use the above equation and
the values of fz/U- 1 shown at the bottom of the table.
4. Application: Use for bare ducts, not furred or insulated. Use the
values for bright metal ducts for both unpainted aluminum and
unpointed galvanized ducts. Condensation at elbows, transformations and other fittings will occur at (1 higher supply air
temperature because of the higher inside film heat transmission
coefficient due to the air impinging against the elbow or fitting.
For low velocity fittings, assume an equivalent velocity of two
U = overall heat transmission coefficient of duct
Etu/(hr) (sq ft) (deg F)
f2 = film heat transmission coefficient on outside of duct, Btu/(hr)
(sq ft) (deg F) = 1.65 for painted ducts and 1.05 for bright
metal ducts.
times the straight run velocity and use the above table. For
higher velocity fittings where straight run velocity is 1500 fpm
and above, keep the supply air temperature no more than one
degree lower than the room dewpoint. Transformations having a
slope less than one in six may be considered as a straight run.
5. Bypass Factor and Fan Heat: The air leaving the dehumidifier
will be higher than the apparatus dewpoint temperature when
the bypass factor is greater than zero. Treat this OS a mixture
problem. Whenever the fan is on the leaving side of the dehumidifier, the supply air temperature is usually at least one to
four degcxr higher than the air leaving the dehumidifier, and
con be calculated using the fan brake horsepower.
6. Dripping: Condensation will generally not be severe enough to
cause dripping unless the surface.temperature
is two to three
degrees below the room dewpoint. Note that the table is based
on the duct surface temperature equal to the room dewpoint
in estimating the possibility of dripping. It is recommended that
the surface temperature be kept above the room
dewpoint.
7. Elimination of Condensation: The supply air temperature must
be high enough to prevent condensation at elbows and fittings.
Occasionally, it might be desirable to insulate only the elbows
or fittings. If moisture is expected to condense only at the fittings,
apply insulation I%” thick usually sufficient) either to the inside
or outside of duct at the fitting and for a distance downstream
equal to 1.5 times the duct perimeter. If condensation occurs on
a straight run, the thickness of insulation required ccm be found
by solving the above equation for U.
f
9o” REDUCING TEE
Air
Duct
Control
In low velocity air distribution systems the flow
of air to the branch take-off is regulated by a splitter
clampcr. The position of the splitter damper is atl,justcd by wise of the splitter rod. Splitter dampers
for rectangular duct systems arc illustrated in Fig.
36. Pivot type dampers are sometimes installed in
the branch line to control Bow. When these are used,
splitter dampers are omitted. Splitter dampers arc
preferred in low velocity systems, and pivot type or
volume dampers are used in high velocity systems.
Condensation
D u c t s m a y “ s w e a t ” when the surface tempera’
ture of the duct is below the clewpoint of the
surrounding air. Table 4 lists the maximum clifference between supply air temperature ant1 room dewpoint without condensing moisture
on the duct for
various duct velocities. See the notes below the table
for application of the data contained in the table.
Table 5 lists various U factors for common insulating material. It can be used in conjunction with
Table 4 to determine required insulation to eliminate
condensation.
In high velocity systems, balancing or volume
dampers are required at the air conditioning terminals to regulate the air quantity.
TABLE 5-DUCT HEAT TRANSMISSION COEFFICIENTS
TYPE DUCT INSULATION
FINISH
Uninsulated Sheet Metal
NOW
Metal lath and ploster-3~”
Wood lath and plaster-%”
Corkboard
NOlV.2
None
Plaster-%”
Plaster--3/s”
2
2.9
-
.12
Corrugated Asbestos Paper (air cell)
None
None
1
7
0.73
1 “‘
SO
.34
311
Rock
NOW3
None
.OE
.17
.27
-
.21
.lO
39
.26
Cork
Plaster-%”
Plaster-%”
Mineral Wool Blanket
None
None
Glass
NOM
1
2
NOIll?
1
05%
Fiber
Magnesia
*Conductivity of insulating material only (per in.)
tOvera U for still air outside duct and 1200 fpm inside duct.
$Uninsulated
Bare Duct.
Air
Velocity (fpm)
Overoll
u
1 400
1 800
1 1200
/
1600
/ 2000
1 .98
1 1.08
/
1
1.19
/
1.14.
1.22
1.0
(:I I,\I”I‘I~l< 2. , \ I I< l)lJ(:‘l‘
Ifb:SI(;N
2-89
tlislril~titiott
DUCT SYSTEM ACCESSORIES
Pit-e th111~~33,
xc
;tc(~cssories
systcttt
;~ucss d o o r s ;ittd sound
wltidt
ttt;ty
1x2
I)itt do riot tti;itcri:tlly
tlwx:
;trc sevctxl
tttenl,
tltc ;ttltlitioti;il
twogtii~ctl
htttjws
;il~ec:t
;iI~sorI~crs
iii 2
tltc tlcsigtt,
iii scrics.
rcsist;ittcc
wlieii sclectittg
t-quit4
t
~Jttdcrwritet3
inst;tll;itiott
duct
‘I‘lterc
itttlcss
;iir Ilow tttttst
I)e
the r:ttI.
Sotw;tlly
21itl
;tt-e
I.
7‘!1e
N;tlioti;tl
tlcscrilx3 rltc:
I)txctic-cs
t
w
gcttetxl
itt Ixttttlhlct
o
Ih~;ttd
01
I’irc
cotlstritc~liott
;iiid
Nh1;IJ
~wittcil~~tl tyjws
!)O,\.
01’ fire tl;rtttlxt3
ductwork:
rcctangttl;tr
iti;iy IX tisetl
pivot tlatttper
t
(I’ig.
in either the vcrtic;tl
I) whit
It
or Ilori/orit;ll
Insitioii.
Fire Dampers
tiott
‘I‘ltc
usctl i t t tu.t;tngrtI;tr
l;or this :trtxttgco
systetn.
loc21
or st;ttc
cotistructioti
cotlcs di’ct;ttc
tltc IISC, loc;t-
01 l i t - c d;itttIxx-s
[or
FUSIBLE
LINK
ANGLE
211
air
L’. T l i e
rec:t:ingular
i)c usctl
loliver
litx d;itttpet-
which
only in tlic horizontal position (Fig.
BAR,
STOP
TRUNNION
BAR
I
(PREFERRED)
I
I
I
I
I
/ BLADE
ANGLE
STOP
SPRING
CATCH
REQ’O FOR BLADES
i
OPEN
I
CONNECTION
I
I
I
_
SLEEVE
CLOSED
/
lMaximum
Over-all Height
Maximum Over-all Width
Minimum Sleeve Length
Sleeve
Blade - Up to 18”
181/,,” - 36”
36%/’ and over
Frame Bearing Support
Trunnion
Bar
Spring Catch
I
I
4
I^
POSITION
MATERI.\L
0
&I
/
~~~~~~~
POSITION
H A T
SPECIFICATIONS
30”
50”
113A”
10 U.S. gage steel
16 U.S. gage steel
12 U.S. gage steel
7 U.S. gage steel
3” x 7/8” x I/B” hat channel
Die cast steel
0.040” bronze spring stock
SECT
A-A
C H A N N
E
L
ntay
JJ).
’ I’;\K-1‘ 2.
2-30
ICig7rve f i illustr;ites
;I l)ivot lirc damper
Access
Doors
;\ccess doors or ;~cccss pnels 3re required in duct
s y s t e m s bclorc ;1nt1 :lftcr ecjuipent installctl i n
ducts. Access panels ;ire dso rquirctl lor access to
l’tlsiblc links in fire thnlpers.
l)lS’I‘IIIl~U’l‘ION
DUCT DESIGN
lor round
duct systcrus. ‘I‘his thnip III:L~ bc used in either the
hori~ont;tl or vertic;il I,osition.
r\lK
I‘liis sectioll presents the ncceswry thta lor ticsig-!litig l o w :tntl high v e l o c i t y duct systetris. This
ht;i inrlutlcs the stanthrcl :lir Iriction chrts, rccoii~tuentlctl tlcsign velocities, losses tllru ell~ows ;~ncl
fittiligs, ;tntl the con1u1on niethotls ol designing the
air distribution systenls. 1nforni;ltion is given also
lor ev:~lu;~ting the effects ol duct hwt glin ;intl ;tltitutle 011 systeni tlesjgn.
- SLEEVE
1” x 1” x L’
AANGLE STOP
.
FUSIBLE LINK BAA
BLADE LINKAGE TIE ROD
(2 REO’D FOR BLADES OVER 38’1
-COUNTERWEIGHT
r FUSISLE LINK
-LINKAGE
AIR
FLOW
(ACCEPTABLI
CLEVIS
(PREFERRED)
ANGLE TIE BAR -
- FRAME BEARING SUPPORT
HAT CHANNEL
- BLADE CONNECTOR
/ HOLDING TAB
-SPRING CATCH
( 2 REO’DFOR
BL&DESOVER 38’)
MATERIAL
Maximum Over-all Height
Maximum Over-all Width
Minimum Sleeve Length
Maximum Blade Width
Sleeve
Blade
Frame Bearing Support
Blade Linkage Roci
Trunnion Bar
Spring Catch
SPECIFICATIONS
91 I/”
10”
1 l!V,”
6”
10 U.S. gage steel
16 ‘J. S. gage steel
3” x 7/g” x l/g” hat channel
B/tu” die CRS
Die cast steel
0.040” bronze spring stock
F1c.44-- KECTANCULAK LOUVER FIKEDAMPF.K
.
ANNEL FRAME
SECTION A-A
o f these f a c t o r s i s illustrated i n t h e f o l l o w i n g
equation:
Friction Chart
In any duct section thru which air is flowing,
tllcrc is a continuous loss of pressure. This loss is
callctl duct friction loss and depends on the Iollowing:
wh&: AP = frjction loss (in. wg)
~~ f= interior surface roughness (0.9 for galvanized duct)
L = length of duct (ft)
d = duct diameter (in.), equivalent diam.
for rectangular ductwork
V = air velocity (fpm)
1. Air velocity
2. Duct size
5. Interior surface roughness
4. Duct length
Varying any one of these four factors inlluences
the friction loss in the ductwork. The relationship
SHEET METAL
ACCESS PANEL
FUSIELE LINK BAR
-_..
-
PREFERRED)
i
(ACCEPTABLE)
BLADE
- SLEEVE
DUCT
( 2 REP D FOR DIMPERS
L
MINIMUM
’
’
ANGLE CLIPS TO FASTEN
SLEEVE IN FIRE PARTITION
OPEN POSITION
MATERIAL
CLOSED POSITION
SPECIFICATIONS
Maximum Diameter
48”
Minimum Sleeve Length
Sleeve
Blade - Up to 18”
18r/,,” to 36”
364/,~” and overt
Trunnion
Bar
Spring Catch
151/2” plus wall thickness*
10 U. S. gage steel
16 U.S. gage steel
12 U.S. gage steel
7 U. S. gage steel
Die cast steel
0.040” bronze spring stock
“.4ccess panel in sleeve. Length 8” plus wall thickness when
access panel is in duct.
tRequires
s” x v,” x l/s” angle blade stiffener.
I;IG. 45 - ROUNLI PIVOT FIRE DAMPER
SECTION A-A
( BEARING
ASSEMBLY)
The u2tut’ll ducts Ior a high velocity supply systeiii
I‘his equation is used to construct the standard
Ii;tve tllc S;IIIIC tlcsigll \,elocity
reroilllllcncl~~tions
as
I’rictiorl chart (Cli,r~‘l 7) IMsctl o n
galv;tiliKtl duct
listed ill I-‘trl~le 7 lor a low velocity systcln, unless
antI air at i0 1; ;IIICI ‘L!).!)‘L ill. I-lg. *l‘his chart nlay I)c
wsctl I’or systciils h;intllirlg ;iir Iroin Y) t
o
I20 I; csteiisive sound trcatnlcnt is provitlctl to iise liighc1
velocities.
;intl lor altitiitlcs ril) t o 2000 It without correcting
tlrc air tleflsity. 1’0clge 5 9 c o n t a i n s tlie data l’or tleFriction Rate
signing high altitude air tlistril)ution systcllls.
‘l‘hc friction rate 011 the Uriction chart is given in
terins
0L inches of w a t e r per 1 0 0 l‘t ol equivalent
Air Quantity
length of duct. To cletcrmine the loss in any section
The total s u p p l y air qiiailtity and t h e q u a n t i t y
of ductwork, the total equivalent length in tilat
required for each spiw is deterniincd from the air
section is multiplictl by the friction rate which gives
contlitioniilg load estiniatc in P0rl /.
the friction loss. The total equivalent length ol duct
includes all elbows and fittings that may IX in the
Duct Diameter
duct section. Tnbles 9 thm 2,3 w e used to cvaluatc
Trrble 6 gives the rectangular duct sizes for the
the losses thrii various duct systein elements in terms
various equivalent duct diameters shown on C1tnl.t 7.
of equivalent length. The duct sections including
Next to the tliametcr is the cross-section area of the
these elements arc measured to the ccnterlinc of the .
round duct. The rectangular ducts shown for this
elbows in the duct section as illustrated in Fig. -Ih.
cross-section area handle the same air quantity at
The fittings are measured as part of the duct section
the same friction rate as the equivalent round duct
having the largest single dimension.
listed. Therefore, this cross-section area is less than
Velocity Pressure
the actual cross-section area of the rectangular duct
The friction chart shows a velocity pressure condetermined by multiplying the duct dimensions. In
version line. The velocity pressure may be obtained
selecting the rectangular duct sizes from Tnble 6,
by reading vertically upward from the intersection
the duct diameter from the friction chart or the
of the conversion line and the desired velocity.
duct area as determined from the air quantity and
TnOle 8 contains velocity pressures Eor ‘the corvelocity may be used.
responding
velocity.
However, rectangular duct sizes should not be
determined directly from the duct area without
Flexible Metal Conduit
using Table 6. If this is done, the resulting duct
Flexible metal conduit is oEten used to transmit
sizes will be smaller, and velocity and friction loss
the air from the riser or branch headers to the air
will be greater, for a given air quantity than the
conditioning terminal in a high velocity system.
design values.
The friction loss thru this conduit is higher than
Air Velocity
The design velocity for an air distribution system
depends primarily on sound level requirements, first
cost and operating cost.
Table 7 lists the recommended velocities for
supply and return ducts in a low velocity system.
These velocities are based on experience.
In high velocity systems, the supply ducts arc
normally limited to a maximum duct velocity Of
5000 fpm. Above this velocity, the sound level may
become objectionable and the operating cost (friction rate) may become excessive. Selecting the duct
velocity, therefore, is a question of economics. A
very high velocity results in smaller ducts and lower
duct material cost but it requires a higher operating
cost and possibly a larger fan motor and a higher
class fan. If a lower duct velocity is used, the ducts
must be larger but the operating cost decreases and
the fan motor and fan class may be smaller.
\
.
thru round duct. Chart 8 gives the friction rate for
3 and 4 in. flexible metal conduit.
(Cotttinzud
on page 38)
OFFSETS
XOTE;
FIG.
*All measurements are center line. Fittings are
measured as part of the duct having largest single
dimension.
46 -
GUIDE
FOR
b~/IEA.SURINC l>uc:-I’
LEN(;.I.HS
(:t-I.\I’-I‘EI<
2. .\II<
I)u(:~I‘
2-33
i)kSl<;N
CHART 7-FRICTION LOSS FOR ROUND DUCT
.02
.03
0 4
.05 0 6
08 0 . 1
0.15
0.2
0.3
0.4 0.5 0.6
0.8
1.0
1.5 2.0
3.0
4.0 5.0 6.0
100
60000
70000
60
000
50000
40000
30000
20000
15 0 0 0
IO 000
IO 0 0 0
JO0
8000
7000
7000
6000
6000
5000
5000
4 0 0 0
5
k
3 0 0 0 ;-’
F
2000
0
LL
z
1 5 0 0
I500
5
s
(I
z
1000
1000
800
700
800
700
600
600
500
5 0 0
4 0 0
4 0 0
80
70
60
50
“” .02
.03
.04
.05 .06
.08
0.1
0.15
0.2
0.3
0.4
0.5 0.6
0.8
I.0
I.5
2.0
FRICTION LOSS (IN. WG PER 100 FT OF EQUIVALENT LENGTH)
3.0
4 . 0 5.0 6.0
I
TABLE
7
SIDE
10
6-DUCT
6
8
1
wea
Diam
q ft in.
DIMENSIONS,
1
SECTION AREA, CIRCULAR
AND DUCT CLASSt
12
14
Area
Diem
*q
ft
in.
4rea
Diam
sq
ft
in.
lo
AleCi
Diam
sq
ft
in.
Art?0
rqft
Diam
in.
.39
.45
8.4
9.8
12
9.1
0.7
.65
.77
10.9
11.9
14
.52
9.8
1.5
.91
12.9
.94’
1.09
EQUIVALENT
16
4rect
1q ft
Diam
in.
DIAMETER,*
18
20
22
ArMI
Diam
sq
ft
in.
Area
Diam
sq
ft
in.
ArMI
Di.¶m
sq
ft
in.
13.1
14.2
1.20
15.3
16
1.24
.I
1.45
16.3
1.67
17.5
18
20
1.40
.O
1 .b3
17.3
1.87
18.5
2.12
19.7
1.54
.8
1.81
18.2
2.07
19.5
2.34
20.7
2.61
21.9
15.9
16.6
1 .b9
17.6
1.99
19.1
2.27
20.4
2.57
21.7
2.86
22.9
3.17
14.6
1.38
1.50
1.83
18.3
2.14
19.8
2.47
21.3
2.78
22.6
3.11
23.9
3.43
24.1
25.11
1.26 /
15.2
1.61
17.2
1.97
19.0
2.31
20.6
2.66
22.1
3.01
23.5
3.35
24.8
3.71
26.1
1.33
15.6
1.71
7.7
2.09
19.6
2.47
21.3
2.86
22.9
3.25
24.4
3.60
25.7
4.00
27.1
13.6
1.41
16.1
1.8
.3
2.22
20.2
2.64
22.0
3.06
3.46
25.2
3.89
26.7
4.27
28.0
14.0
1.48
16.5
1.93 a8.8
2.36
20.8
2.81
22.7
3.25
23.7
24.4
3.68
26.0
4.12
27.5
4.55
28.9
14.4
1.58
17.0
2.03
2.49
21.4
2.96
23.3
3.43
25.1
26.7
4.37
28.3
4.81
29.7
2.61
21.9
3.11
23.9
3.63
25.6
3.89
4.09
29.0
5.07
30.5
::::
2.76
22.5
3.27
24.5
3.80
26.4
4.30
27.4
28.1
4.58
:f:; I ::;:
4.84
29.8
5.37
31.4
40
2.88
23.0
3.43
25.1
3.97
27.1
4.52
28.8
5.07
30.5
5.62
32.1
42
44
2.98
23.4
4.15
4.33
27.t
28.:
29.4
23.9
25.6
26.1
4.71
3.11
3.57
3.71
4.90
30.0
5.31
5.55
31.2
31.9
5.86
6.12
32.8
33.5
46
3.22
24.3
3.88
24.8
4.49
4.65
28.2
29.:
30.6
3.35
26.7
27.2
5.10
48
50
5.30
31.2
5.76
5.97
32.5
33.1
6.37
6.64
34.2
34.9
22
24
.78
12.0
1.08
14.1
.84
12.4
1.16
26
.89
12.8
28
.95
13.2
30
32
.Ol
I.07
34
I.13
36
I.18
38
1.23 4
19.3
II%-%
3.46
25.2
4.03
4.15
27.6
5.51
31.8
6.19
33.7
6.87
35.5
52
2.22
20.2 2 . 9 1
23.1
3.57
25.6
430
28.1
5.72
32.4
6.41
34.3
7.14
36.0
54
2.29
20.5
2.98
23.4
28.5
5.17
30.1
5.90
32.9
6.64
34.9
7.38
36.8
2.38
20.9 3.09
23.8
26.1
26.5
4.43
56
3.71
3.83
4.55
28.9
5.31
31.:
6.08
33.4
6.87
35.5
7.62
37.4
29.3
5.48
31.;
6.26
33.9
7.06
36.0
7.87
38.0
58
60
3.94
4.06
26.9
27.3
4.68
4.84
29.8
5.65
32.:
6.50
34.5
7.26
36.5
8.12
38.6
64
4.24
27.9
5.10
30.6
5.91
33..
6.87
35.5
7.71
37.6
8.59
39.7
68
4.49
28.7
5.37
31.4
6.26
33.!
7.18
36.3
8.12
38.6
9.03
40.7
72
76
4.71
29.4
5.69
4.91
30.0.
5.86
32.3
32.8
6.60
6.83
34.q
35..
7.54
7.95
37.2
38.2
8.50
8.90
39.5
40.4
5.17
30.8
6.15
5.41
31.5
7.22
7.54
36..
37..
8.55
39.0
39.6
9.21
9.75
88
5.58
32.0
6.41
6.64
33.6
34.5
8.29
84
34.9
7.87
38.1
8.94
40.5
10.1
92
5.79
32.6
38.t
9.39
41.5
10.4
43.8
5.90
33.0
35.6
36.2
8.12
96
100
6.91
7.14
36.9
8.40
8.50
39..
39..
9.70
9.80
42.1
7.40
10.8
11.3
44.5
45.5
12.1
12.3
47.2
47.6
104
7.60
37.4
8.90
40..
10.3
43.5
11.6
46.2
13.0
48.8
108
7.90
112
8.10
38.0
38.6
9.20
9.50
41.
41 .I
10.6
10.9
44.0
44.7
12.0
12.3
47.0
47.5
13.4
13.8
49.6
50.3
9.80
10.0
42.
42.E
11.3
48.1
49.1
14.3
51.3
13.1
14.4
51.5
10.3
43.5
.5
0 .o
6
11.9
46.7
12.6
11.
13.4
49.6
15.0
52.4
10.6
44.1
12.1
47.1
13.8
50.4
15.5
53.3
12.5
12.8
47.9
48.5
14.1
14.5
50.9
51.6
15.8
16.2
53.9
54.5
48.8
49.4
14.7
15.2
52.0
52.9
lb.5
55.0
55.6
4.15
80
27.6
116
120
124
128
I
I
132
136
140
144
L
*Circular equivalent diameter (d,). Calculated from d, = 1.3 $b)
10.4
43.8
10.8
44.6
X’
13.0
13.3
b) “’
M
I Large numbers in table ore duct class.
42.5
,
lb.8
.
(:f f.\l”f‘El<
2. :\IK DUC’1’ I~BSIGN
2-35
TABLE 6. DUCT DIMENSIONS, SECTION AREA, CIRCULAR EQUIVALENT DIAMETER.*
AND DUCT CLASS? (Cont.)
24
26
AWXI
Diem
s q f t in.
Alea
*q ft
28
Diam A r e a
in.
rqft
36
32
Diam
in.
4rea
'4 ft
Diam
in.
4rea
sq ft
Diam
in.
Al=30
Diam
s q f t in.
30
AK70
Dism
s q f t in.
40
AlfK!
sq ft
Diam
in.
ArMI
Diam
sq ft in.
16
18
20
22
24
3.74
26.2
26
4.03
27.2 4.40
4.33
28.2 4.74
29.5 5.10
4.68
29.3
5.07
30.5 5.44
30.6
31.6
5.86
32.8
4.94
30.1
5.37
31.4 5.79
32.6
6.23
33.8
5.68
5.24
33.6
6.60
b.99
34.8
7.06
36.0
7.54
37.2
35.8
36.7
7.46
37.0
7.95
38.2
8.46
7.87
38.0
8.37
39.2
8.89
a.29
39.0
8.81
40.2
9.34
0.68
9.21
9.61
41 .l
10.1
10.5
28.4
36
5.58
31 .o
32.0
5.69
5.94
32.3 6.15
33.0 6.51
3a
5.86
32.8
6.38
3C.2 6.07
40
6.15
33.6
6.71
35.1 7.22
42
6.45
34.4
7.03
35.9 7.58
36.4
37.3
44
6.75
35.2
7.34
36.7 7.91
38.1
8.50
39.5
9.07
39.9
40.8
46
7.03
35.9
37.4 a.25
36.6
38.2 8.59
8.85
9.25
40.3
41.2
41.7
7.30
7.58
38.9
39.7
9.48
48
50
7.63
7.95
37.3
a.25
38.9. 0.90
40.4
9.61
42.0
9.89
10.3
42.6
43.5
52
54
7.87
38.0
a.55
39.6 9.25
41.2
9.98
11.4
38.7
8.85
40.3 9.61
42.0
10.4
10.7
11.0
44.3
a.16
a.42
42.8
43.6
45.0
11.8
39.3
9.16
41.0 9.94
42.7
10.7
44.3
11.4
45.8
58
60
8.63
39.8
9.48
41.7 1 0 . 3
43.4
11.0
a.a9
9.75
42.3 10.5
44.0
11.4
11.8
12.2
64
9.43
40.4
41.6
45.0
45.8
46.6
10.3
43.5 11.2
45.4
12.1
47.2
56
68
9.98
72
10.4
42.8
43.8
76
10.8
44.9
80
84
--
11.5
12.0
46.0
-
_
34.6
35.5
7.34
9.43
41.6
9.89
42.6
9.80
41.4
42.4
10.4
43.6
10.5
11.0
43.8
44.8
42.0
10.3
43.4
10.8
44.6
11.4
45.8
43.0
10.7
11.3
45.6
43.9
44.8
11.1
11.6
44.3
45.2
46.5
47.4
46.8
47.8
46.1
11.8
12.2
11.9
12.4
13.0
48.8
45.7
46.5
12.1
12.6
47.1
12.7
13.5
49.7
48.0
13.2
48.3
49.2
12.2
47.3
13.0
48.8
13.7
50.1
Es?
12.6
48.1
13.4
49.6
14.2
51.0
15.0
52.4
47.3
13.0
48.9
13.8
50.4
14.6
51.8
12.9
48.7
13.8
50.4
14.7
52.0
15.5
53.4
15.5
16.5
53.3
55.0
14.6
51.8
56.6
58.0
10.9
10.9
44.7 11.8
46.6
12.8
48.4
13.7
50.2
15.6
53.5
16.5
55.0
17.5
11.5
12.0
45.9 12.4
47.0 13.1
47.8
13.5
49.7
14.4
51.5
16.4
17.4
56.5
49.0
14.1
50.8
15.1
52.7
17.3
54.9
56.3
18.3
57.9
18.3
19.3
12.6
13.2
48.0 13.7
49.2 14.2
50.1
14.7
52.0
15.8
53.9
17.0
55.8
18.1
46.9
17.3
17.7
18.5
57.0
58.2
18.9
19.7
59.3
60.7
50.1 1 4 . 8
55.0
56.3
19,2
20.1
13.7
53.2
54.3
57.6
58.9
47.9
15.4
16.1
16.5
12.5
51.1
52.2
60.1
20.9
62.0
12.9
48.7
14.2
51.1 15.5
53.4
16.7
55.4
18.0
57.4
19.2
59.4
20.5
61.3
21.8
63.2
56.2
57.3
la.6
19.2
58.5
59.4
19.7
20.6
60.2
61.5
21.1
62.2
22.7
64.5
21.6
63.0
23.4
65.5
60.5
21.4
62.6
22.7
64.5
24.1
66.5
61.4
22.0
63.5
23.5
65.7
24.8
67.5
c .6 2 . 3
22.5
64.3
24.5
67.0
25.7
96
13.3
49.5
14.8
52.2 15.9
54.0
17.2
100
13.9
50.6
15.0
52.5 16.7
55.3
17.9
104
14.6
51.8
15.8
53.9 17.1
56.0
18.6
108
14.8
52.1
16.2
54.6 17.6
56.8
19.2
58.5
59.4
19.9
20.5
112
15.1
52.7
16.8
55.5 18.3
58.0
19.7
60.1
21.1
116
15.8
120
124
16.2
16.6
128
17.1
17.4
56.0
132
136
*Circular
39.4
40.4
61 .O
21.2
62.4
2
2
23.0
.
1
65.0
24.0
24.8
66.3
67.5
25.6
26.5
68.5
69.7
68.7
27.1
70.5
26.2
69.4
20.2
71.9
27.2
28.2
70.6
71.9
29.0
73.0
74.0
56.4 la.9
.9
20.3
61 .l
22.0
63.6
23.5
65.7
24.8
17.8
18.4
57.1
19.4 o
.6
5 8 . 1 1 9 . 8 c 3. 3
20.9
62.0
22.7
64.5
21.6
63.0
23.2
65.4
24.2
25.2
66.7
68.0
26.1
26.5
la.8
19.3
58.8 20.3
22.3
22.6
64.0
64.4
25.6
26.3
68.6
69.5
27.3
28.2
72.0
28.7
29.8
72.6
74.0
30.2
24.5
66.0
67.0
70.8
59.5 20.8
61.1
61.8
23.7
56.5
32.0
74.5
76.6
17.9
57.3
19.7
60.2 21.4
62.7
23.0
65.0
25.1
67.9
26.9
70.3
28.7
72.6
30.5
74.8
32.6
77.3
18.5
58.2
20.3
61.0
22.3
64.0
24.1
66.5
25.9
69.0
27.5
71.1
29.4
73.5
31.5
76.0
33.4
78.3
la.8
58.7
20.6
61.5 22.7
64.5
24.8
.Gr.;
67.5
26.3
69.5
28.2
72.0
29.9
74.1
32.0
76.6
34.0
79.0
55.2
diameter
(d,.).
Calculated
from
d,
=
1.3
2 lb)
T
ttorge numbers in table are duct class.
67.5
69.2
20.3
17.3
equivalent
53.9
54.6
59.5
69.8
29.8
,
2-06
i’,\R’I‘ 2. i\IK I)IS-I‘KIBU?‘ION
TABLE 6. DUCT DIMENSIONS, SECTION AREA, CIRCULAR EQUIVALENT DIAMETER,*
AND DUCT CLASS1 (Cont.)
.
I
60
I
64
I
68
I
I
72
76
42
44
46
48
50
52
54
56
58
60
23.5
65.7
64
25.0
67.7
26.7
70.0
28.3
72.1
30.2
74.4
31.8
33.5
76.4
78.4
26.5
69.7
72
28.0
76
129.5
74.1
71.7 29.9
73.6 I 31.q6.1
80
31.0
75.4
33.
84
32.5
77.2
34.8
68
88
96
*Circular
34.0
137.0
79.0
82.4
equivalent
36.3
139.8
diameter
78.8
80.9
37.7
83.2
8.1 135.2
79.9 37.0
80.4 I 37.4
82.8
39.6
85.3
82.4
39.2
84.8
41.4
81.6
84.2
41.1
86.8
43.4
87.2
89.3
47.5
93.4
85.5
(d,.).
142.1
Colculoted
.
87.9/44.6
from
I
II
I
/
33.8
35.7
18.6
I
d,.
=
90.5/
1.3 (q + b).‘”
41.7
ttarge
I
87.5
numbers in table
ore ducr &~I.
I
I
I
(:I-I.\l”l‘~:K
2.
,\lli
I)lJ(:‘I’
DI3I(;N
237
TABLE 7-RECOMMENDED MAXIMUM DUCT VELOCITIES FOR LOW VELOCITY SYSTEMS (FPM)
CONTROLLING
CONTROLLING FACTOR
NOISE GENERATION
Main Ducts
APPLICATION
FACTOR-DUCT
Main Ducts
I
FRICTION
Branch
Ducts
Return
SUPPlY
SUPPlY
Return
600
1000
800
600
600
Average Stores
Cafeterias
1800
2000
1500
1600
1200
Industrial
2500
3000
1800
2200
1500
Residences
Theatres
Auditoriums
General
Offices
High Class Restaurants
High Class Stores
Banks
TABLE 8-VELOCITY PRESSURES
VELOCITY
PRESSURE
(in. wg)
.Ol
.02
.03
.04
.OS
.06
.07
08
.09
.lO
.11
.12
.13
.14
.15
.16
.17
.lB
.19
.20
.21
.22
.23
.24
.25
.26
.27
.2a
VELOCITY
(Ft/Min)
400
565
695
800
895
980
1060
1130
1200
1270
1330
1390
1440
1500
1550
1600
1650
1700
1740
1790
1830
1880
1920
1960
2000
2040
2080
2120
VELOCITY
PRESSURE
(in. wg)
.29
.30
.31
.32
.33
.34
.35
.36
.37
.38
VELOCITY
(Ft/Min)
2150
2190
2230
2260
2300
2330
2370
2400
2440
2470
I
.39
.40
.41
.42
.46
.47
.48
, __ __
2500
2530
2560
2590
I
2710
2740
2770
.53
.54
.55
.56
NOTES: 1. Data for standard air (ZP.W in. Hg and 70
2. Data derived from the following equation:
2910
2940
2970
2990
I
I
VELOCITY
PRESSURE
(in. wg)
.58
.60
.62
.64
.66
.68
.70
.72
.74
.76
.7a
.BO
.a2
.a4
.86
.68
.90
.92
.94
.96
.9a
1.00
1.04
1.08
1.12
1.16
1.20
1.24
VELOCITY
(Ft/Min)
I
I
3050
3100
3150
3200
3250
3300
3350
3390
3440
3490
3530
3580
3620
3670
3710
3750
3790
3840
3880
3920
3960
4000
4080
4160
4230
4310
4380
4460
I
I
I
VELOCITY
PRESSURE
(in. wg.)
1.28
1.32
1.36
1.40 1.44
1.48
1.52
1.56
1.60
1.64
1.68
1.72
1.76
1.80
1.84
1.88
1.92
1.96
2.00
2.04
2.08
2.12
2.16
2.20
2.24
2.28
VELOCITY
(Ft/MinI
.I
A530
I
I
I
I
4600
4670
4730
4800
A870
49%
5000
5060
5120
5190
5250
5310
5370
5430
5490
5550
5600
5660
5710
5770
5830
5080
-I-5940
i690
6040
FJ
where: V = velocity in fpm.
h, = pressure difference termed “velocity head” (in. wg!.
CHART 8-PRESSURE DROP THRU
FLEXIBLE CONDUIT
FAN CONVERSION l.OSS OR GAIN
In addition to the calculations shown for deter-
mining the required static pressure at the fan discharge in l<xnrtzple -t, a fan conversion loss or gain
must be included. This conversion quantity can be
a significant amount, particularly on a high velocity
system. It is determined by the following equations.
if the velocity in the duct is higher than the fan
outlet velocity, use the following lormuIa for the
additional static pressure required:
Loss = I.][(&)”
-(&)‘]
where T’,t = duct velocity
V, = fan outlet velocity
Loss = in. wg
If the fan discharge velocity is higher than the
duct velocity, use the following formula for the
credit taken to the static pressure required:
Cain = 35 [(,&>l - (‘$&)L’]
DUCT SYSTEM ELEMENT FRICTION LOSS
Friction loss thru any fitting is cspt~essetl in terms
of equivalent length of duct. This method provides
units that can be used with the friction chart to
tletertnitic the loss in ;I section of duct containing
elbows and fittings. ToOk l-7 gives the friction losses
for rectangular elbows, and TuOle /I gives the losses
for standard round elbows. The friction losses in
TU/ll6% If o,rd 12 arc given in terttts of additional
equivalent length of straight duct. This loss for the
elbow is added to the straight run of duct to obtain
the total equivalent lctigth of duct. ‘l‘lie
straight run
.
of duct is measured to the intersection of the center
line of the fittings. Fig. 46 gives the guides for mcasuring duct lengths.
Tables 9 and 10 list the friction losses for other
size elbows or other R/D ratios. Table 10 presents
the friction losses of rectangular elbows and elbow
combinations in terms of L/D. Table 10 also includes the losses and regains for various duct shapes,
entrances and exits, and items located in the air
stream of the duct. This loss or regain is expressed
in the number of velocity heads and is represented
by “n”. This loss or regain may be converted into
equivalent length of duct by the equation at the
end of the table and added or subtracted from the
actua1 duct length.
Table 9 gives the loss of round elbows in terms of
L/D, the additional equivalent length to the diam- .
eter of the elbow. The loss for round tees and crosses
are in terms of the number of velocity heads (“n”).
The equation for converting the loss in velocity
head to additional equivalent length of duct is
located at the bottom of the table.
In high velocity systems it is often desirable to
have the pressure drop in round elbows, tees, and
crosses in inches of water. These losses may be obtained from Chart 9 for standard round fittings.
DESIGN.,
METHODS
The general procedure for designing any duct system is to keep the layout as simple as possible and
make the duct runs symmetrical. Supply terminals
are located to provide proper room air distribution
(Chapter j), and ducts are laid out to connect these
outlets. The ductwork should be located to avoid
structural members and equipment.
The design of a low velocity supply air system
may be accomplished by any one of the three following methods:
1. Velocity reduction
2. Equal friction
3. Static regain
The three methods result in diBerent levels of
accuracy, economy and use.
The equal friction meti~od is rccommencled
return and exhaust air systems.
for
LOW VELOCITY DUCT SYSTEMS
Velocity Reduction Method
The procedure for designing the duct system by
this method is to select a starting velocity at the
fan discharge and make arbitrary reductions in
velocity down the duct run. The starting velocity
selected sl~oulcl not exceed those in Table 7. Equiva-
CHAPTER 2. AlIt DIJCT DESIGN
2-39
TABLE O-FRICTION OF ROUND DUCT SYSTEM ELEMENTS
ELEMENT
CONDITION
L/D
RATIO’
R/D = 1.5
9
90’ 5-Piece Elbow
R/D = 1.5
12
45’ j-Piece Elbow
\
R/D = 1.5
6
45’ Smooth Elbow
R/D = 1.5
90’ Smooth Elbow
90’ j-Piece Elbow
i
fv-.;/
90’ Miter Elbow
r-l
ELEMENT
90’ Tae$ and 90°,
135O
B 180”
Voned
Not Vaned
Thru
22
65
CONDITION
VALUE OF n f
0.2
0.5
Fl- :*;
1 .
4.0
Crossf
v2 _
Pressure Loss
4.5
I
ga
B r a n c h = nhv,
.lO
44
i.21
1.47
Pressure Loss
Thru
B r a n c h = nhvs
90’ Conical Tee and 180”
Conical Cross
v2
-=
Vl
N o t e s o n page
42.
I.
0.5
1.0
g
0.2
0.5
1.0
1.2
.;g$
PAKT 2. AIR DISTRIRUTION
2-m
TABLE IO-FRICTION OF RECTANGULAR DUCT SYSTEM ELEMENTS
CONDITIONS
ELEMENT
Rectangular Radius
Elbow
.5
W/D
.5
1
3
Voned
X0 Elbow
No
Rectangular Square Elbow
Elbow
Double
9
11
14
18
5
7
8
‘12
j
1
4
4
5
7
18
12
10
Radius Elbow
10
8
7
8
7
7
7
7
6
:/PO times value for
imilar 90’ elbow
60
Single Thickness Turning Vanes
15
Double Thickness Turning Vanes
‘0.
s=o
15
S=D
10
22
15
Elbow
S=D
1.25* For Both
Elbow
Elbow
14
18
30
40
Vanes
Elbow
W/D = 2. RI/D
1 1.50
1 .oo
20
W/D = 1, R/D = 1.25*
Double
or Unvaried
S = D
W/D = 1. R/D, = 1.25*
W/D = 1, R/D =
.75
Elbow
1
2
3
Double
/
33
45
80
125
6
Double
1.25*
t
L/D R a t i o
hc
’
QY!
Double
RATIO
R/D
r
P
Rectangular Vonad Radius
1 L/D
16
Direction of Arrow
= 1.25*, RZ/D
= .5
Reverse Direction
w
W / D = 4. R/D = 1.25* for both elbows
Direction of Arrow
Reverse Direction
18
l
TABLE lo-FRICTIONOF RECTANGULAR DUCT SYSTEM ELEMENTS (Contd)
ELEMENT
CONDITIONS
Transformer
V A L U E O F nl
I
.15
v:! = Vl
S . P . Loss = n h v ,
’
Expansion
5O
A3
.89
.93
~ZPI
.20
.40
.bO
“il”
Angle “a”
15O
10”
.74
33
.87
S.P. Regain = n(hvt
20”
.b0
.78
.a4
.b2
.74
A2
30°
.52
40”
.45
.b8
.79
.b4
.77
-
a
1
3o”
1
45’
1
60”
n
1
1.02tt
1
1.04
1
1.07
S . P . L o s s = n(hvl -hvl) ttslope 1 0 i n 4“
Abrupt
.35
Entrance
S . P . Loss = n h v ,
Bellmouth
Abrupt
Entrance
Exit
-4
S.P. Loss or Regain Considered Zero
I
Bellmouth
Re-Entrant
Exit
Entrance
.
S.P. Loss = nhv,
Sharp Edge Round Orifice
0
*I
-*2
.85
A,/A,
1
0
1
”
1
2.5
)
.25
2.3
1
)
SO
1.9
1
1
.75
1.1
S.P. Loss = nhvl
Abrupt
Contraction
v,/v~
”
1
1
1
1
1.00
0
T
-5
1
1
0
1.34
.25
1.24
1
1
.50
.9b
1
1
.75
.52
S.P. Loss = nhv?
Abrupt
Expansion
V~/VI
n
1
1
1
1
.20
.32
.40
.48
.bO
.48
1
1
1
1
S . P . R e g a i n = nhv,
Pipe Running Thru
Duct
E/D
n
1
I
.lO
1
.25
1
.50
.20
I
.55
I
2.00
.lO
/
.25
I
.50
.7
1
1.4
I
4.00
S.P. Loss = nhv,
Bar
Running Thru Duct
E/D
”
1
1
S.P. Loss = nhv,
Easement
Over
Obstruction
f-1
E/D
n
1
.I0 1 .25
.07 1 .23
S.P. Loss = nhv,
Notes on page 42.
1
1
.50
.90
.80
.32
NOTES FOR TABLE 9
NOTES FOR TABLE 10:
*L and D are in feet. 0 is the elbow diameter. L is the additional
equivalent length of duct added to the measured length. The
equivalent length L equals D in feet times the ratio listed.
tThe value of n is the loss in velocity heads and may be converted
to additional equivalent length of duct by the following equation.
$1.25 is standard for on unvoned full radius elbow.
tl and D are in feet. D is the duct dimension illustrated in ths
drawing. L is the additional equivalent length of duct added to the
measured duct. The equivalent length 1 equals D in feet times
the ratio listed.
Lcnxhv;flOO
where: 1
=
h, =
.
tThe value n is the number of velocity heads or differences in
velocity heads lost or gained at a fitting, and may be converted
to additional equivalent length of duct by the following equationr
additional equivalent length, ft
velocity pressure at Vz, in. wg (conversion line on
Chart 7 or Table 8).
h t = f r i c t i o n loss/100
(Chart 7).
f t , d u c t d i a m e t e r a t VZ, i n . wg
where: L
hv
=
additional
=
velocity pressure for VI or V2,
line on Chart 7 or Table 8).
” = value for tee or cross
equivalent
length,
hf = friction loss/100 ft. duct
(Chart 7).
ITee or cross may be either reduced or the some size in the straight
thru portion
n
=
cross
ft.
in. wg
section at
(conversion
h,, in. wg
value for particular fitting.
.
TABLE 11-FRICTION OF ROUND ELBOWS
90’ S M O O T H
R / D = 1.5
90’ 5-PIECE
R/D =
1.5
A D D I T I O N A L EGUIVALENT
3
4
5
3
4
5
6
2.3
3
3.8
4.5
7
5.3
8
6
7
:
9
10
11
12
14
16
18
20
22
24
6
10
-
-
90° 3-PIECE
R/D =
6
8
10
12
14
lb
18
20
11
12
14
22
24
20
32
18
20
22
24
36
lb
1.5
I
LENGTH OF STRAIGHT DUCT
40
44
48
.’
45’ O - P I E C E
45’ S M O O T H
R / D = 1.5
R / D = 1.5
(FT)
.
1.5
2
2.5
3
1.1
1.5
1.9
2.3
3.5
4
4.5
5
2.6
5.5
6
7
0
9
10
11
12
3
(
I
(:1l,\l”l‘lcI~
II.
.\lII
l)li(:
I‘
2-40
Ill-SIGN
TABLE 12-FRICTION OF RECTANGULAR ELBOWS
RADIUS ELBOW
NO VANES
DUCT
W
D
RADIUS
Radius Ratio+
R/D = 1.25
ELBOW-WITH VANES1
R t = 6”
(Recommended)
SQUARE ELEOWSf
I
Rt = 3”
(Acceptable)
Double
Thickness
Turning Vanes
Single
Thickness
Turning Vanes
ADDITIONAL EQUIVALENT LENGTH OF STRAIGHT DUCT (FT)
96
45
36
31
33
28
2
2
2
1
43
31
38
29
25
3
3
2
2
2
40
30
25
20
17
60
45
37
30
25
28
23
21
1715
13
12'
44
33
28
29
23
18
2
2
2
1
1
1
41
29
33
25
19
16
15
3
3
2
2
2
2
1
35
29
25
21
18
15
11
60
45
37
30
2
20
15
48
36
30
24
20
16
12
27
22
19
16
14
12
10
41
31
25
27
22
16
2
2
2
1
1
1
39
27
31
26
21
15.
14
3
3
2
2
2
2
1
33
27
23
20
17
13
10
60
45
37
30
25
20
15
96*
48
36
30
24
20
16
12
10
a
45
26
20
la
15
14
11
9
a
a
35
35
26
23
24
19
15
3
2
2
2
1
1
1
34
22
28
21
17
14
13
11
9
3
3
2
2
2
2
1
1
1
29
23
21
ia
15
12
10
a
7
60
A5
37
30
25
20
15
12
10
42
36
30
24
20
16
12
10
23
20
17
15
13
11
9
a
28
24
21
21
la
14
2
2
2
1
1
1
26
21
26
19
16
13
13
10
8
3
3
2
2
2
2
1
1
I
24
22
20
16
14
12
9
a
6
53
45
37
30
25
20
15
12
10
36
72*
36
30
24
34
19
16
14
27
22
19
20
3
2
2
1
34
ii.
'ix
1 37
1
19
22
22
15
12
12
9
a
3
2
2
2
2
1
1
1
20
la
15
13
11
9
a
6
45
37
30
25
20
15
12
10
19
ia
19
1612 -
2
2
1
1
1
16
21
17
14
12
12
9
a
3
2
2
2
2
1
1
1
17
17
15
12
11
a
7
6
40
37
30
25
20
15
12
10
72
60
48
42
~._
48
36
30
24
31
25
22
19
48
36
30
.24.
20
16
12
a I
10
a
32
32
30
24
20
16
12 10
a
I
I
*
a
7
17
16
14
-l2
10
a,
7
6,
1
s
5
/
TABLE 12-FRICTION OF RECTANGULAR ELBOWS (CONT.)
DUCT
DIMENSIONS
(in.)
RADIUS ELBOW
NO VANES
RADIUS
ELBOW-WITH
VANES1
SQUARE ELBOWS$
R’
W
Radius Ratio+
R/D = 1.25
D
R, = 3”
(Acceptable)
R+ = 6”
(Recommended)
Double
Thickness
Turning Vanes
Single
Thickness
Turning Vanes
ADDITIONAL EQUIVALENT LENGTH OF STRAIGHT DUCT (FT)
Vfltle5
28
28
24
20
16
12
10
8
24
20
15
13
12
10
a
14
17
15
11
2
1
1
1
17
15
13
11
11
9
8
7
6
96*
72%
40*
24
20
16
12
10
8
6
38
32
22
19
17
20
lb
13
11
3
3
2
1
1
1
80*
60*
40*
20
16
12
10
8
6
32
26
16
19
15
12
9
3
2
2
1
1
64*
48*
32*
16
26
21
15
8
5
4
48*
36*
24*
12
10
19
16
11
7
6
5
4
16
v
6
20
14
12
10
10
8
7
3
2
2
1
9
12
11
8
Vall.3
$J
2
2
2
2
1
1
I
3
2
2
2
1
1
1
14
10
9
9
8
7
3
2
2
;1
1
1
12
9
8
0
6
6
3
3
2
1
1
1
14
13
12
10
0
7
6
34
30
25
20
15
12
10
23
21
18
12
10
9
8
7
6
4
80
72
62
30
25
20
15
12
10
8
19
17
14
10
8
7
6
5
4
66
58
14
12
11
7
6
5
5
4
48
43
38
20
15
12
10
8
.
49
25
20
15
12
10
0
8
2
2
1
8
7
8
7
5
5
3
3
2
1
1
1
10
9
8
5
5
4
3
33
30
26
15
12
10
8
6
2
2
1
6
8
6
5
3
2
2
1
6
19
13
9
5
4
4
8
7
6
4
4
3
27
24
21
12
10
8
32*
24*
16'
8
6
13
11
8
4
3
5
6
4
2
1
1
21
19
16
10
6
24%
18*
12*
6
10
4
3
1
1
15
13
11
8
12
a
6
10
40*
30*.
20*
10
a
7
8
8
6
I
3
t F o ro t h e r r a d i u s r a t i o s , s e e T a b l e I O .
- D e n o t e sH a r d B e n d s a s s h o w n
Hard Bend
3
Easy Bend
.
J\' I6
I'
.
I F o ro t h e r s i z e s , s e e T a b l e1 0 .
V a n e s m u s t b e l o c a t e d a s i l l u s t r a t e d i n C h a r t6, page 24, to
have these minimum losses.
I
;
I
CHART 9-LOSSES
FOR ROUND FITTINGS
Elbows, Tees and Crosses
NOTES: 1. Loss for tee or cross is CI function of the velocity in the branch. This represents the loss in static pressure from the main upstream
to the branch. Qlr is the ratio of air quantity of the branch to the main upstream.
2. Loss for
45’ smooth elbow is
equal to one-half the loss for a 90’ smooth elbow.
3. Loss for 45’ 3-piece elbow is equal to one-half the loss for a 90’ 5.piece
elbow.
2-46
~ I’,IliT 2 . A I R DISTIIIBUTION
lent round diameters may be obtained from Cllnrl i
using air velocity and air quantity. Tab/e 6 is usctl
with the equivnlcnt round diameter to select the
rectangular duct sizes. The fan static pressure rcquiretl for the supply is detcrminctl by calc~llation,
using the longest run of duct including all elbows
and tittings. TnOle.5 /I) ~161 12 are used to obtain the
losses thru the rectangular elbows and littings. The
longest run is not necessarily the run with the
greatest friction loss, as shorter runs may have more
elbows, fittings and restrictions.
This method is not normally used, as it requires
a broad background of duct design experience and
knowledge to be within reasonable accuracy. It
should be used only for the most simple layouts.
Splitter dampers should be included for balancing
purposes.
metrical layouts. II’ a design has a mixture of short
and long rumi, the shortest run requires consicleral)le tlampering. Such a system is difficult to balance
since the equal friction method makes no provision
for cqnali/:ing pressure drops in branches or for
providing the same static pressure behind each air
terminal.
The usual I)rocedure is to select an initial velocity
in the main duct near the fan. This velocity should
be selcctetl from Tcrble 7 with sound level being the
limiting factor. Clrnrt 7 is ~csctl with this initial
velocity and air quantity to determine the friction
rate. This same friction loss is then maintained
th~~ghout the system and the equivalent round
duct tliametcr is selected from Clrt~~.t 7.
Equal Friction Method
To expedite equal friction calculations, Table 13
is often used instead of the friction chart; this results
in the same duct sizes.
Tliis method oE sizing is used for supply, exhaust
and return air duct systems and employs the sa,me
friction loss per foot of length for the entire system.
The equal friction method is superior to velocity
reduction since it requires less balancing for sym-
The duct areas determrned from Table 13 or the
equivalent round diameters from C/NW 7 are used
to select the rectangular duct sizes from Table ci.
This procedure of sizing duct automatically reduces
the air velocity in the clirection of flow.
TABLE
CFM
CAPACITY
%
I~-PERCENT
DUCT
AREA
%
SECTION
PREA
IN
BRANCHES FOR,MAINTAINING
/
EQUAL
CFM /
CAPACITY
%
DUCT
AREA
%
CFM j
CAPACITY
%
DUCT
AREA
%
CFM /
CAPACAY
%
FRICTION
1
2
3
4
5
2.0
3.5
5.5
7.0
9.0
26
27
28
29
30
33.5
34.5
35.5
36.5
37.5
51
52
53
54
55
59.0
60.0
61.0
62.0
63.0
76
77
78
79
80
DUCT
AREA
%
81.0
82.0
83.0
84.0
84.5
6
7
8
9
10
10.5
11.5
13.0
14.5
16.5
31
32
33
34
35
39.0
40.0
41.0,
42.0
43.0
56
57
58
59
60
64.0
65.0
65.5
66.5
67.5
81
82
83
84
85
85.5
86.0
87.0
87.5
88.5
11
12
13
14
15
17.5
18.5
19.5
20.5
21.5
36
37
38
39
40
44.0
45.0
46.0
47.0
48.0
61
62
63
64
65
68.0
69.0
70.0
71.0
71.5
86
87
88
89
90
89.5
90.0
90.5
91.5
92.0
16
17
18
19
20
23.0
24.0
25.0
26.0
27.0
41
42
43
44
45
49.0
50.0
51.0
52.0
53.0
66
67
68
69
70
72.5
73.5
74.5
75.5
76.5
91
92
93
94
95
93.0
94.0
94.5
95.0
96.0
24
25
31.5
32.5
59
-50
57.0
58.0
74
75
80.0
80.5
99
100
99.0
100.0
’
l
CN/\I’TER 2 . AIR
247
VTSIGN
7’0 determine the total
sYste*ll that the fan m u s t 0, ,oss in t,,e t,Lict
to dwhte t h e l o s s
i n the ‘G,,.t is necessary
highest resistance. The friction 1, havi,,l: the
and fittings in the section must be ,L,all ei,,ows
Example 4 - Equal Friction Method of &“},
‘\
Given:
Duct systems for general office (rig. 47).
Total air quan’ity - 5400 cfm
18 air terminals - 300 cfm each
Operating pressure for all tcrlninals - 0.15
Radius ell)ows, R/D = 1.25
‘?lCfS
‘_.
DUCT
SECTION
1
in, wg
Solution:
1. From
P)
.4DD.
EQUIV
ENCTH
’
(f9
60
Duct
A- B
” B - 1 3
-
Duct
Elbow
Duct
Duct
T&/r
7 selecx an initial velocity of 1700 fpm.
;;>I
Duct
17 _ 18\‘.Uct
cfm
5400
Duct area = s = 3.18 sq ft
- From
\
\
Find:
1. Initial duct velocity, area, sire and friction r a t e in fhp
duct section from the fan to the hrst branch.
2. Size of remaining duct runs.
3. Total equivalent length of duct run with highest resistance.
otal sratic pressure required at fan discharge.
I
LENGTH
ITEM
Table 6,selec; a duct size - 22 in. x 22 in.
Initial friction rate is determined from Chart 7 using
the air quantity (5400), and the equivalent round duct
diameter from Table 6. Equivalent round duct diameter
= 24.1 in.
i
Friction rate = ,145 in. wg
2. The duct areas are
sizes are determined
lat.% thh design information:
DUCT
SECTION
AIR
QUANTITY
CFM*
CAPACITY
(cfm)
(%)
5400
3600
100 \
ToA
A-B
B- 13
13-14
14-15
15-16
1800
1580
1200
900
16 - 17
17 - 18
600
300
I
DUCT
SECTION
ToA
A-B
B- 13
13 - 14
14 - 15
15 - 16
16. 17
17 - 18
(%)
69 4
3.18
loo.0
73.5 7’
2.43
1.3
41.0
35.5
29.5
24.0
1.12
.94
.76
.56
.33
17.5
10.5
i
Nair
6
DUCT
SIZE:
(in.)
22x22
22x 16
22x 10
18 x
14 x
12 x
8x
8x
10
10
10
10
10
quantity in duct section
total air quantitv
tDuct area = percent of area times initial duct area (fan
to A)
:Refer to page 21 for reducing duct sizes.
y&!‘/.[f-
The total frictioni
\<he ductwork from the fan to
last terminal I8 is sho
LOSS = total equiv length.ihe f”‘lowing’
‘&tion r a t e
= 229ftx .I45 in. wg i
100ft
=&Jo, i .33 in. wg
Total static pressure required?&.
‘an discharge is the
sum of the terminalLoperating_pr~~
eye and the loss in
the ductwork. Credit can be taken fo
r% velocity regain
between the first and last sections of ducq
11
I
AREA+
: a; t‘hA
4.
210
67
33 28
22
17
DUCT
AREA
*Percent of cfm =
Total
\
f+...’
J”, (
5400 CFM
$3
900
C2-ti 9 0 0
/ 4
22
10
600
16
(26’ 6 0 0
I4
:20
;jJ,,,
1
,2dF
I
12
26 9 0 0 :bC
-, ..’
26 600.
,,
300
j, -s.
r,
‘,
&- ,-
_-1..
2d
3&
/
‘6,
Frc.47 - DUCT LAYOUTFOR Low VEI.OCITY SYSTEM
(EXAMPLES~ AND 4)
Velocity in initial section = 1700 Cpm
Velocity in last section = 590 fpm
ITsin a 75’,‘{, regain coellicient,
= .75 (.I8 - AX) = .I’)N in. wg
Thcrcfore,
the total static pressure at fan discharge:
= duct friction + terminal pressure - regain
= .33 + .I5 - .I2
= .36 in. wg
The equal friction method does not satisfy the
design criteria of uniform static pressure at all
branches and air terminals. To obtain the proper
air quantity at the beginning of each branch, it is
necessary to include a splitter damper to regulate
the flow to the branch. It may also be necessary to
ave a control device (vanes, volume damper, or
.djustable terminal volume control) to regulate the
flow at each terminal for proper air distribution.
In Example 4, if the fan selected has a discharge
velocity of 2000 fpm, the net credit to the total static
pressure required is determined as described under
“Fan Conuersion Loss or Gain:”
Gain = .75 [(gy
-($j&r]
= .75 (.25 - .lS) = .05 in. wg
Static Regain Method
The basic principle of the static regain method is
to size a duct run so that the increase in static pressure (regain due to reduction in velocity) at each
branch or air terminal just offsets the friction loss
in the succeeding section of duct. The static presqure is then the same before each terminal and at
lch branch.
The following procedure is used to design a duct
system by this method: Select a starting velocity at
the fan discharge from Table 7 and size the initial
duct section from Table 6.
The remaining sections of duct are sized from
Chart 10 (L/Q Ratio) and Chart II (Low Velocity
Static Regain). Chart 10 is used to determine the
L/Q ratio knowing the air quantity (Q) and length
(L) between outlets or branches in the duct section
to be sized by static regain. This length (L) is the
equivalent length between the outlets or branches,
including elbows, except transformations. The effect
of the transformation section is accounted for in
“Chart II - Static Regain.” This assumes that the
transformation section is laid out according to the
recommendation presented in this chapter.
Chart ff is used to determine the velocity in the
duct section that is being sized. The values oE the
L/Qrntio(Chnrt
IO) and the velocity (V,) in the duct
section immediately before the one being sized are
used in Chart Il. The velocity (VJ determined from
Chart II is used with the air quantity to arrive at
the duct area. This duct arcx is used in Table 6 to
size the rectangular duct and to obtain the equivnlent round duct size. By using this duct size, the
friction loss thru the length of cluct equals the increase in static pressure due to the velocity change
after each branch take-off and outlet. However,
there are instances when the reduction in area is too
small to warrant a change in duct size after the
outlet, or possibly when the duct area is reduced
more than is called for. This gives a gain or loss for
the particular duct section that the fan must handle.
Normally, this loss or gain is small and, in most instances, can be neglected.
Instead of designing a duct system for zero gain
or loss, it is possible to design for a constant loss or
gain thru all or part of the system. Designing for a
constant loss increases operating cost and balancing
time and may increase the fan motor size. Although
not normally recommended, sizing for a constant
.
loss reduces the duct size.
Example 3, - Static Regain Method of Designing Ducts
Given:
Duct layout (Example 4 and Fig. 47)
Total air quantity - 5400 cfm
Velocity in initial duct section - 1700 fpm (Example 4)
Unvaned radius elbow, R/D = 1.25
18 air terminals - 300 cfm each
Operating pressure for all terminals - 0.15 in. wg
Find:
1. Duct sizes.
2. Total static pressure required at fan discharge.
Solution:
1. Using an initial velocity of 1700 fpm and knowing the
air quantity (5400 cfm), the initial duct area after the
fan discharge equals 3.18 sq ft. From Table 6, a duct
size of 22” x 22” is selected. The equivalent round duct
size from Table 6 is 24.1 in. and the friction rate from
Chart 7 is 0.145 in. wg per 100 ft of equivalent length.
The equivalent length of duct from the fan discharge
to the first branch:
= duct length i- additional length due to fittings
= 60 + 12 = 72 ft
The friction loss in the duct section up to the
branch:
= equiv length of duct X friction rate
0.145
=72x100=
first
.104 in. wg
The remaining duct sections are now sized.
The longest duct run (A to outlet IS, Fig. 47) should be
sized first. In this example, it is desirable to have the
i
.
249
CHAPTER 2. AIR DUCT DESIGN
CHART IO--L/QRATIO
.OS
.04
.03
I
.01
a6
-I 0
13
1
’
2
II’
3
“““’
4 567810
I!
20
1.8,
I,,
30 40 5060 80 100
AIR O”ANT,TY AFTER TAKE-OFF, 0 t 100 CFMI
From Form E-147A
CHART II-LOW VELOCITY STATIC REGAIN
VELOCITY AFTER TAKE-OFF,V2
( FPM)
From Form E-147A
2750
I’.ART 2 . A I R DLSTRII~UTION
static prcsswe
in the duct immediately before outlets f
and 7 equal 10 the static pressure Iwfore outlet 13.
is a savings in building space normally allotted to
the air conditioning ducts.
Figure -48 tal~ulatcs the duct sizes.
Usually Class 11 fans are rcquirctl for the increased static pressure in a high velocity system and
extra care must be taken in duct layout and construction. Ducts are normally sealed to prevent leakage of air which may cause objectionable noise.
Round ducts are preferred to rectnngular because of
greater rigidity. Spirn-Pipe
should be used whenever
possible, since it is made of lighter gage metal than
corresponding round and rectangular ducts, and
does not require bracing.
2. The total pressure required at the fan discharge is
equal to the sum of the friction loss in the initial tlucl
section plus the terminal operating pressure.
Fan
discharge pressure:
= friction loss + terminal pressure
= ,104 + .I5 = 2.5 in. wg
It is good design practice to include splitter
dampers to regulate the flow to the branches, even
though the static pressure at each terminal is nearly
equal.
Comparison of Static Regain and Equal Friction Methods
Examples 4 and 5 show that the header duct sizes
determined by the equal friction or static regain
method are the same. However, the branch, ducts,
ized by static regain, are larger than the branch
ducts sized by equal friction.
Figure 49 shows a comparison of duct sizes and
weights established by the two methods.
The weight of sheet metal required for the system designed by static regain is approximately 13%
more than the system designed by equal friction.
However, this increase in first cost is offset by reduced balancing time and operating cost.
If it is assumed that a low velocity air handling
system is used in Examples 3 and 4 and that a design air flow of 5400 cfm requires a static pressure
of 1.5 in. wg, the increased horsepower required for
other equal friction design is determined in the following manner.
Symmetry is a very important consideration when
designing a duct system. Maintaining as symmetrical
a system as possible recluccs balancing time, design
time and layout. Using the maximum amount of
symmetrical duct runs also reduces construction and
installation costs.
Particular care must be given to the selection and
location of fittings to avoid excessive pressure drops
and possible noise problems. Figure 50 illustrates
the minimum distance of six duct diameters between elbows and 90” tees. If a 90” conical tee is
used, the next fitting in the direction of air flow may
be located a minimum of one-half duct diameter
away (Fig. 51). The use of a conical tee is l&ted to
header ductwork and then only for increased initial
velocitiei in the riser.
When laying out the header ductwork for a high
velocity system, there are certain factori that must
be considered:
1. The design friction losses from the fan discharge to a point immediately upstream of the
first riser take-off from each branch header
should be as nearly equal as possible. These
points of the same friction loss are shown in
Fig. 52.
Additional hp = 1*861y51*75
= 6.3’% approx.
A 6% increase in horsepower often indicates a
larger fan motor and subsequent increased electrical transmission costs.
HIGH
VELOCITY
DUCT
SYSTEMS
A high velocity air distribution system uses higher
air velocities and static pressures than a conventional system. The design of a high velocity system
involves a compromise between reduced duct sizes
and higher fan horsepower. The reduced duct size
2. To satisfy the above principle when applied to
multiple headers leaving the fan, and to take
maximum advantage of allowable high velocity, adhere to the following basic rule wherever
possible: Make as nearly equal as possible the
ratio of the total equivalent length of each
header run (fan discharge to the first riser
take-off) to the initial header diameter (L/D
ratio). Thus the longest Spiru-Pipe header run
should preferably have the highest air quantity
s o that the highest velocities can be used
throughout.
3. Unless space conditions dictate otherwise, the
take-off from the header should be made using
a 90” tee or 90” conical tee rather ‘than a 45”
Cr-l;\I”l‘EK
2-51
2. AIK I>U(:‘I’ D E S I G N
rota1 S.P. Loss for
SUI3~1~
.
duct system = S.P. for critical duct _ in. wg plus air outlet S.1’. loss - in. wg = _ in.
3
4
I<()UlV.
I
Q
1
5
L
I
1
VELOCITY
1 Indicated
I
6
7
ARE,4
DIJCT
DI.\M.
OR
RECT.
srzlq
1
(in.)
Fan to A
.A - B
B -13
13 - 14
14 - 15
5400
3600
1800
1500
1200
22 x
22x
22x
22x
22x
15 - 16
16 - 17
17 - 18
B- 7
7- 8
900
600
300
1800
1500
20x 10
16 x 10
10 x 10
22x 10
22x 10
5- 9
-10
iO- 11
ll- 12
A- 1
1200
900
600
300
1800
22x 10
20 x 10
16 x 10
10 x 10
22 x 10
1500
1200
900
600
300
22x 10
22x10
20x 10
16 x 10
10 x 10
l2345.
2
3
4
5
6
D U C T
SECTIOF
48
- DUCT
S IZING
C ALCULATION
FORA<
Duct
Dimensions
(in.)
Duct
Weight
(‘b)
592
179
394
411
2-3,8-g, 14-15
3-4, 9-10, 15-16
4-5, 10-11, 16.17
5-6, 11-12, 17-18
14 x
12 x
8x
8x
10
10
10
10
360
321
270
270
Total weight of duct*
Allow 15y0 for scrap
Total wt of sheet metal
22 x
22x
22x
22x
’
2797
420
3217
transformation
and
I
Duct
Dimensions.
(in.)
22
16
10
10
includes
0.104
STATIC REGAIN METHOD
22 x
22x
22x
18x
weight
0.104
EQUAL FRICTION METHOD
ToA
A to B
A-l, B-7, B-13
l-2, 7-8, 13-14
*Total
“‘. 2
16
10
10
10
From Form E-147
not symmetrical and handle different air quantities, an
initial velocity is assumed at the beginning of the branch.
This velocity is somewhat less than the velocity in the
header before take-off.
*Duct size is assumed to determine loss thru elbow.
tDuct sizes from Table 6. Longest duct run is sized first.
Remaining duct sections are the same size, as they are
, symmetrical to branch B tkm 18. If other branches are
F IG.
9
1 F R I C T I O N 1TOT‘\L
LOSS OR
S.P.
TAKE-OFF
LOSS
TO
IN
TAKE-OFF. DUCT
S.P.
CHANGE
(in. wg)
(in. wg)
1 Selected 1 Indicated j Selected
(cfm)
8
Wg.
elbows.
F IG. 49 - COMPARISON
OF
DUCT S I Z I N G M E T H O D S
Duct
Weight
(lb)
22
16
10
10
592
179
394
438
22x 1 0
20 x 10
16 x 10
1 0 x 10
438
435
384
297
3157
475
3632
2-52
I’,\RT
tee. 1Sy using 90” fittings, the pressure drop to
the branch throughout the system is more uniform. In addition, two fittings are normally
required when a 45” tee is used and only one
when a 90” fitting is used, resulting in lower
first cost.
The design of a high velocity system is basically
the same as a low velocity duct system ciesigneci for
static regain. The air velocity is reduced at each
take-off to the riser and air terminals. This reduction in velocity results in a recovery of static pressure (velocity regain) which offsets the friction loss
in the succeeding duct section.
The initial starting velocity in the supply header
depends on the number of hours of operation. TO
achieve an economic balance between first cost and
90’TEES
FIG .
50
- S PACING
OF
FITTINGS
IN
DUCT RUN
operating cost, lower air velocities in the header
are recommended for 24-hour operation where space
permits. When a 90” conical tee is used instead of a
90” tee for the header to branch take-off, a higher
initial starting velocity in the branch is recommended. The following table suggests initial veiocities for header and branch duct sizing:
RECOMMENDED INITIAL VELOCITIES
USED WITH CHARTS 12 AND 13 (fpm)
HEADER
12 hr. operation
24 hr. operation
3000 - 4000
2000 - 3500
BRANCH*
90” conical tee
90”
tee
-
4000 - 5000
3500 - 4000
TAKE-OFFS
TO
TERMINALS
5
1 - SPAC:ING
OF
FI-JXNGS
2000
maximum
*Branches are defined as a branch header or riser having 4
to 5 or more take-offs to terminals.
Static regain charts are presented for the design of
high velocity systems. Chart 12 is used for designing
branches and Chart 13 is used for header design.
The basic difference in the two charts is the air
quantity for the duct sections.
Chart 12 is used for sizing risers and branch
headers handling 6000 cfm or less. The chart is
based on 12 ft increments between take-offs to the
air terminals in the branches or take-offs to the risers
in branch headers. A scale is provided to correct for 1
spacings more or less than 12 ft.
Chart 13 is used to size headers, and has an air
quantity range of 1000 to 40,000 cfm. The chart is
based on 20 ft increments between branches. A correction scale at the top of the chart is used when
take-off to branch is more or less than 20 ft.
Examples 6 and 7 are presented to illustrate the
use of these two charts. Example 6 is a branch sizing
problem for the duct layout in Fig. 53, and Example 7 is a header layout (Fig. 55).
90’ CONICAL TEE -,
FIG .
2 . AIR DIS?‘RIBUTION
WHEN USING
90"
h+&
CONICAL TEE
.
2-53
CHAPTER 2. AIR DUCT DESIGN
CHART 12-BRANCH HIGH VELOCITY STATIC REGAIN
R O U N D DUCT EOUIVALENT
LENGTH FOR L * 12 FT
.
.REFERENCE
LINE L
=
12 FT
BRANCH
TAKE-OFF
LOSSI IN WC
DUCT DIAMETER
(IN
,
Form E-l 48A
2-54
PAKT 2. AIK DISTRIBUTION
CHART 13-HEADER HIGH VELOCITY STATIC REGAIN
- - R E F E R E N C E L I N E L = 2OFT
60 55
50
3.0
45
40
36
32
26 26 24
22
20
16
;
16
14
12
I’
II
IO
’
”
,,,,I,!
0
55
50
45
40
36
32
9
’
8
’
I,
7
I
”
i/I
i,
/
/I I II
26 26 24 2
OUCT DIAMETER (IN.)
Form E-149A
CHAPTER
2.
AIR
DUCT
2 - s
DESIGN
CONICAL
90° TEE
SMOOTH
t
SECT- I 1200 CFM
A.T. - 1.5 IN. ug
SMOOTH
90’ ELL
S E C T - 2 IOOOCFM
*
POINT IN BRANCH
HEADER BEFORE
A.T. . 1.5 IN. wg
12’
SECT-3
800
CFM
A.T. = 1.5 IN. vrg
12’
SECT-4 600 CFM
12’
SECT-5
A.T. 8 1.5 IN. w(l
MAIN HEADER
400
CFM
A.T. = 1.5 IN. wg
12’
SECT-6 200 CFM
A.T. = 1.5 IN. wg
k IG. 52 - HIGH V ELOCITY HEADERS
Example 6 - Use of Branch Duct Sizing
.
FIG.
BRANCHES
AND
FOR
E XAMPLE 6
2. Enter CAnrt 12 at the velocity range recommended for
branch risers with a 90” conical tee, Page 52.
Chart
Given:
Office building riser as shown in Fig. 23
12 air terminals - 100 cfm each .
Total air quantity - 1200 cfm
.\ir terminal static pressure - 1.5 in. wg
Find:
Duct sizes for Sections 1 thru 6, Fig. J3.
Solution:
1. Sketch branch as shown in Fig. 23.
values in columns 2, 3 and 8, Fig. 51.
53 - BRANCH DUCT
3.
Intersect the initial branch air quantity, 1200 cfm, as
shown at point A. Read 7 in. duct size and 3.8 in. wg
loss per 100 ft of equivalent pipe and 4500 fpm velocity.
Enter these values on the high velocity calculations
form (Fig. 54).
4.
From point A, determine header take-off loss by projecting horizontally to the left of point A and read
1.25 in. wg.
5. Enter 1.25 in. wg in Fig. 5f for section 1.
Enter appropriate
6. Determine equivalent length from the header to the
first air terminal take-off:
INITIAL CONDITIONS: Cfm 1200; Duct Size 7 in.; Velocity 4500 fpm.
4
1
I
5
PRESSURE READING
(in. wg)
Initial
Selected
I
6
7
8
TAKE-OFF
TO
TAKE-OFF
S.P.
CHANGE
(4 minus 5)
(in. wg)
S.P.
AHEAD
OF
TAKE-OFF
AIR
TERMINAL
PRESS.
(in. wg)
(in. wg)
Branch T a k e - O f f F. L. = 1.25
Duct Friction Loss
= .89
1.0
1.25
- 0.25
0.84
0.84
0.0
0.57
0.47
+ 0.1
0.32
0.40
- 0.085
0.26
0.24
+ 0.02
Maximum S. P . i s at Section 2 :
f
(in.)
@Pm)
2.14
I .5
7
4500
2.39
2.39
2.29
2.37
2.35
1.5
1.5
1.5
1.5
I .5
7
7
7
6
5
3700
3050
2300
2050
1475
2.39
+
F IG. 54 - HIGH V ELOCITY BRANCH S IZING CALCULATIONS
1.5
I
+ .19
= 4.08
From Form E-148
PAKT 2. AIR DISTRIBUTION
256
Length of pipe = 6 + 12 = 18 ft. One 7 in. smooth ell
= 5.3 ft. Total equivalent length = 18 + 5.3 = 23.3 ft.
Pressure drop = 23.3 X 3.8/ 100 = .89 in. wg.
7. Determine duct size for section 2:
From point ,4 on Ctzart 12, project thru points B and C
to the 1000 cfm line at point D.
8. Determine equivalent length for section 2:
Actual duct length = 12 + 2 = 14 ft. Two smooth 90 ells
= 2 X 5.3 = 10.6 ft. Total equivalent length = 14 f 10.6
= 24.6 ft.
9. Determine pressure loss in section 2:
Project vertically from point D to reference line, then
to point E. Proceed on the guide lines to 24.6 ft equivalent
length, point F. Project vertically from I; to 1000 cfm
line at point C, then along the 1000 cfm line to point H.
Enter point H (1.25 in. wg) and point G (1.0 in. wg) in
Fig. 54, columns 4 and 5. The net loss is “point H point C” = 1.25 - 1.00 = .25 in. wg. This is entered in
column G of Fig. 54. Enter 7 in. diameter in column 9.
Example 7 -
Use of Header Sizing Chart
Given:
Office building, 12.hour
operation
Header as shown in Fig. 55
10 branches - 1200 cfm each
Total air quantity - 12,000 cfm
Find:
Header size for sections 1 thru 10
Solution:
I. Sketch header as shown in Fig. 55.
values in Fig. 56, columns 1, 2, 3 and 8.
Enter
appropriate
2.
Enter Chart 13 at the velocity range recommended
headers in a system operating 12 hours, page 5~‘.
3.
Intersect the initial header air quantity 12,000 cfm as
shown, at point A. Read 24 in. duct size and .62 in.
wg loss per 100 ft of equivalent pipe and 3800 fpm.
Enter these values on high velocity calculation form
(Fig. 56).
10. Determine duct size for section 3:
Project downward on the 7 in. diameter line to the 1000
cfm line, points H to I.
4.
Calculate the equivalent length of section 1 and record
in column 3; straight duct = 20 feet, no fittings; pressure*
drop = 20 X .62 = ,124 in. wg.
11. Project along guide lines at the right side of the chart
from I to the 800 cfm line at point J. Duct size is 7 in.
Enter appropriate values from the chart in columns 4,
5, 6 and 9 of Fig. 54.
5. Size duct section 2: From point A on chart, project thru
points B and C to the 10,800 cfm line at point D.
12. Determine duct size for section 4:
Project downward on the 7 in. diameter line to the 800
cfm line, points J to K.
13. Project along guide lines at the right side of the chart
from point I< to the 600 cfm line at point L. Project
along the 600 cfm line to the 7 in. diameter line, point
L to M. This results in a static regain of .57 - .47 = .I0
in. wg. Duct size for section 4 is 7 in. Enter appropriate
values in Fig. 54, columns 4,5, 6, 7 and 9.
NOTE: If the 600 cfm line is projected from point L to
the 6 in. diameter line, a net loss of .88 - .45
= .43 in. wg results. This friction loss unnecessarily penalizes the system. Therefore, the projection from L is made to the 7 in. diameter line.
14.
Determine duct size for section 5: Project downward
from M to 600 cfm line, point N. Project along guide
lines to 400 cfm line, point 0. Continue along the 400
cfm line to the 6 in. diameter line, point 0 to P. This
results in a static pressure loss of .40 - .315 = .085 in.
wg. Duct size is 6 in. Enter the appropriate values in
Fig. 54, columns 4, 5, 6, 7 and 9.
for
6. Determine equivalent length for section 2: Actual length
= 20 ft. One j-piece 90” ell = 24 ft. Total equivalent
length = 20 + 24 = 44 ft.
7. Determine pressure loss in section 2: Project vertically ,
from point D to reference line, point E. Proceed on the .
guide lines to 44 ft equivalent length, point F. Project
vertically from F to 10,800 cfm line at point G, then
along the 10,800 cfm line to point H . Enter the net
loss., read from point G (.84) and from point H (.90) in
columns 4 and 5, Fig. 56. The net loss is “point H - point
G” = .90 - .84 = .06 in. wg. This is entered in column
6, Fig. 56. Enter 24 in. diameter in column 9.
8. Determine duct size for section 3: Project downward on
the 24 in. line to the 10,800 cfm line, point H to I.
Project along the guide lines at the right side of the
chart from I to the 9600 cfm line at point J. Enter
appropriate values from the chart in columns 4, 5, 6 and .
9, Fig. 56.
NOTE: If the 400 cfm is projected from point 0 to
the 7 in. diameter line, a net regain of .315 20 = ,115 in. wg results. Therefore, the 6 in.
size is used to save on first cost since the net loss
using the 6 in. size is insignificant.
.-
15. Determine duct size for section 6:
Duct size is 5 in. as determined-from point S.
16. Determine velocities for duct sections l-6 from points
A, I, K, N. Q and T respectively;, enter in column 10.
17. Determine take-off and runout
pressure drop by entering upper right hand portion of Chart 12 at 100 cfm and
read a pressure drop of .19 in. wg for a 4 in. runout size.
18. Add 2.39 in. wg (maximum from column 7) plus.1.5 in.
wg (column 8) plus .I9 (take-off and runout drop) to
find 4.08 in. wg (total branch S.P.).
FIG.
55
- HIGH VELOCITY D UCT SYSTEM - HEADER
STATIC REGAIN M ETHOD SIZING
j i
CIl.2I’TER
2.
AIK
DUCT
2-257
DESIGN
INITIAL CONDITIONS: Cfm 12,000; Duct Size 24 in.; Velocity 3800 fpm.
4
1
2
3
5
HEADER
SECT.
NO.
AIR
QUANTITY
Q
EQUIV.
DUCT
LENGTH
L
(cfm)
w
12000
10800
9600
8400
7200
20
44
20
20
20
0.84
0.74
0.57
0.42
6000
4800
3600
2400
1200
44
20
20
20
20
0.31
0.22
0.195
0.165
0.165
PRESSURE
READING
(in. wg)
Initial
Selected
8
9
10
S.P.
AHEAD
OF
TAKE-OFF
BRANCH
S.P.
DUCT
SIZE
VELOCITY
V
(in. wg)
(in. wg)
(in.)
@pm)
6
TAKE-OFF
TO
TAKE-OFF
S.P. CHANGE
(4 minus 5)
(in. wg)
Duct Friction = 0.124
0.90
-0.06
0.70
$0.04
0.55
$0.02
0.42
0.0
0.30
0.26
0.23
0.24
0.2 1
10.01
-0.04
-0.035
-0.075
-0.045
0.124
0.184
0.144
0.124
0.124
4.08
4.08
4.08
4.08
4.08
24
24
24
24
24
3800
3400
3000
2600
2250
0.114
0.154
0.189
0.264
0.309
4.08
4.08
4.08
4.08
4.08
24
22
20
16
12
1900
1800
1650
1650
1500
Maximum S.P. at Section 10 = 0.31
f
4.08 = 4.39
From Form E-149
FIG. 56
-
HIGH V ELOCITY HEADER S IZING CALCULATIONS
9. Determine duct sizes for sections 4 thru 10 in a manner
similar to Step 8, using the listed air quantities and
equivalent lengths. One exception is duct section 6. Since
its equivalent length is 44 feet, use the method outlined
in Steps 5, 6 and 7 to determine the pressure drop. In
addition, see Exampk 5, Steps 13 and 14, for explanation
when the chart indicates a duct diameter other than
those listed, for instance 23 inches.
&CT HEAT GAIN AND AIR LEAKAGE
Whenever the air inside the duct system is at a
temperature different than the air surrounding the
duct, heat flows in or out of the duct. As the load
is calculated, an allowance is made for this heat
gain or loss. In addition, air leakage is also included
in the calculated load. The load allowance required
and guides to conditions under which an allowance
should be made for both heat gain or loss and duct
leakage are included in Part I, System Heat Gain.
Chart 14 is used to determine the temperature
rise or drop for bare duct that has an aspect ratio
of 2: 1. In addition, correction factors for other
aspect ratios and insulated duct are given in the
notes to the chart.
Example 8 - Calculations for Supply Duct
Given:
Supply air quantity from load estimate form - 1650 cfm
Supply duct heat gain from load estimate form - 57,
Supply duct leakage from load estimate form - 5%
Unconditioned space temp - 95 F
Room air temperature - 78 F
Duct insulation U value - .24
Duct shown in Fig. 57.
Find:
Air quantities at each outlet
Solution:
1. Room air quantity required at 60 F
1650
= 1 + .05 + .05 = 1500 cfm
FIG. 57
- DUCT HEAT GAIN
AND
AIR
L EAKAGE
2-58
PART 2. AIR DISTRIBUTION
Normally a 10 c;/o leakage allowance is used if the complete duct is outside the room. Since a large portion of
the duct is within the room. 5(;;IS used in this example.
2. Determine the temperature rise from A to f1: Select an
initial starting velocity from Tab/e 7 (assume 1400 fpm).
Calculate the temperature rise from the fan to the room.
Enter Cltczrt If at 1500 cfm; project vertically to 1400
fpm and read .25 degrees temperature c h a n g e per 100
ft per degree F tlillerence.
Using aspect ratio of 2: I,
temperature
rise
Outler
=522-(1540X&) =492&n
3. Determine cfm for outlet C: Use equal friction method
to determine velocity in second section of duct, with
1540 - 492 = 10-18 cfm; velocity = 1280 fpm.
Dctcrmine temperature rise at outlet: From C/tart 14,
read 32 for 1280 fpm and 1040 cfm. Temperature rise
= 32 X 17.2 X i’& = .83 F
30 ft
= 100 X .27 F change X ,185 X (95 - 60) = 52 F
Supply air temperature tliff = 17.2 - .8 = 16.4 F
Outlet cfm adjusted for temperature rise
:\ir temperature entering room = GO.52 F
Actual air quantity entering room
=
II cfm with allowance for duct cooling
18
= 500 X 1~.4 = 550 cfm
Allowance for duct cooling
78 - GO
7s - GO.52 X 1500 = 1540 cfm
=550-(1048X&)=498cfm
Air temperature rise from A to B
4. Determine cfm for outlet D:
Use equal friction method to determine velocity in third ’
section of duct with 1048 - 498 = 550 cfm; velocity =
1180 fpm.
Determine temperature rise at outlet:
From CI~art IJ, read .43 F for 1180 fpm and 550 cfm.
Temperature
rise
= & X 17.48 X .2i = .33 F
Supply air temperature cliff to outlet B
= i8 - (60.32 + .33) = 17.15 F
Required air quantity to outlet B
18
= 500 x - = 522 cfm
17.2
= .43 X 16.4 X -& = 1.06 F
with no allowance for cooling from the duct.
Supply air temperature diff = 16.4 - 1.1 = 15.3 F
.
CHART 14-DUCT HEAT GAIN OR LOSS
300
4 0 0 500 600
6 0 0 1000
CFM AIR
2000
NOTES:
3000
4000
6000
6ooo 10000
Aspect Ratio Correction
1. with
Baseda on
duct
2:l bare rectangular
ratio.
aspect
Aspect Ratio
1 Round
1 1:l ]3:1
\ 4:l
Correction
1
1.92
j 1.18)
2. If duct is furred-in or insulated, use the following correction factors:
.83
1 1.1
1 5:l
1 6:l
/ 7:l
1 8:l
I 911
/ 10:1
1.26 / 1.35 1 1.43 j 1.5 / 1.58 1 1.65
Furred-in duct
- .45
I n s u l a t e d (U = .27) - ,185
Insulated (U = .13) - .lO
3. For air quantities greater than 10,000 cf m, divide air quantity by 100 and multiply degree change by 0. 1
CHAI’TEK
2.
AIR
DUCT
259
DESIGN
Outlet cfm adjusted for tcmperaturc
HIGH ALTITUDE DUCT DESIGN
rise
18
= 500 x __ = 5X8 cfm
15.3
Allowance for duct cooling
=i88-(i8ax &) =54Gcfm.
5. Check for total cfm:
492 + 498 + 546 = 1536 cfm
This compares favorably with the 1540 cfm entering room.
Fig. 57
shows original and
corrected outlet air quantities.
CHART 15-AIR
0
- 50
2000
0
4000
50
When an air distribution system is designed to
operate above 2000 feet altitude, below 30 F, or
above 120 F temperature, the friction factor obtained from Clrart 7, page 33 using the actual air
quantity at final conditions must be corrected for
air density. Chart 15 presents correction factors for
temperature and altitude. The factors are multiplied together when a system is at high altitude and
also operates outside the temperature range.
DENSITY CORRECTION FACTORS
6000
6000
ALTITUDE (FT)
100
I50
AIR TEMPERATURE (Fl
10000
12000
200
250
14000
300
I’AKT 2. AIK DISTRIBUTION
2-60
DUCT
Tables 15 md Jh which apply for low and high pressure systems. Fig. 58 illustrates the more common
seams and joints used in low pressure systems.
CONSTRUCTION
The sheet metal gage used in the ducts and the
reinforcing required depends on the pressure conditions of the system. There is also a wide variety
of joints and seams used to form the ducts which
also depend on pressure conditions in the duct system.
TABLE 16-MATERIAL GAGE FOR SPIRA-PIPE
DUCT
Low
and High Pressure Systems
Low Pressure Systems
Table 14 lists the recommended construction for
rectangular ducts made of aluminum or steel. The
method of bracing and reinforcing and types of
joints and seams are included in the table. Round
duct and Spira-Pipe construction are included in
TABLE 14-RECOMMENDED
CONSTRUCTION FOR RECTANGULAR SHEET METAL DUCTS
.
Low Pressure Systems
MATER
DUCT
DIMENSION
(in.)
L GAGE
I
Aluminum
B 8 S Gage
Steel
U.S. Gage
RECOMMENDED
CONSTRUCTION*
Transverse Joints, Bracing and Reinforcing
I
Duct
24
Slip
24
Duct
1
Slip
Up to 24
22
I
20
24 t o 3 0
24
24
22
20
3 1 to 60
22
22
20
18
6 1 to 72
20
20
18
16
Reinforced pocket slip? or reinforced Bar-St, spaced not more than four feet
apart.
1 H” x 1 l/i” x x” dio’gonal angle reinforcing$ or 1 I/211 x 1 ‘/a” x gf’ girth
angle reinforcing1 located midway between joints.
16
Reinforced pocket slipt or reinforced Bar-S slip+ spaced not more than four
feet apart.
1 l/i” x 1 l/5” x %” diagonal angle reinforcing? or 1 ‘/a” x 1 %fl x l/e girth
angle reinforcingt located midway between joints.
1 Ya” x ‘/.” band iron stay bracing for duct width 73” to 90”.
16
Reinforced pocket slipt or reinforced Bar-S slipt spaced not more than four
feet apart.
1 y2” x 1 lh” x H” diagonal angle reinforcingf or 1 %” x I’%” x %” girth
angle reinforcing$ lo&d midway between joints.
1 l/q” x 1/” band iron stay bracing for duct width 91 M to 120”.
11%” x I/” b a n d i r o n s t a y b r a c i n g s p a c e d 4 8 ” a p a r t f o r d u c t w i d t h s
121” and uo.
73 to 90
20
91 and Up
20
18
20
18
16
I
i
Pocket slio or Bar-S slio. soaced not more than eight feet aoart.
Pocket slip or Bar-S slip, spaced not more than four feet apart.
I
I
*All ducts over 18” in either dimension ore cross-broken, except those to which rigid board insulation is applied or area of duct where outlet
or duct connection is to be installed. Duct seams are either Pittsburg lock seam or longitudinal seam.
IReinforce joint with 1 l/q” x I/l” band iron.
IAngles
ore attached to duct by tack welding, sheet metal screws, or rivets on 6” centers.
TABLE 15-RECOMMENDED
CONSTRUCTION FOR ROUND SHEET METAL DUCT
Low and High Pressure Systems
MATERIAL
DUCT
DIMENSION
Steel
(in.)
Up to 8
9 to 24
I
RECOMMENDED
GAGE
U.S. Gage
B 8 S Gage
24
22
22
20
25 to 36
20
CONSTRUCTION
Aluminum
18
3 7 to 48
20
18
4 9 to 72
18
16
7 3 and Up
16
14
Reinforcing
I
. Joints and Seams
I
1 l/q” x 1 l/q” x ‘/(” girth angle reinforcing spaced o n 8’ centers.
1 fi” x 1 l/q” x %” girth angle reinforcina soaced on 6’ centers.’
1 IA” x 1 ‘A” x ‘A” girth angle reinforcing spaced on 4’ centers.
Round duct sections ore ioined together by welding, by a coupling,
or by belling out one end of duct.
T h e seams o n r o u n d d u c t mov b e
continuous welded or grooved longitudinal seam.
’
A- DRIVE SLIP
o _ REINFORCED
EAR-S SLIP
FIG. 58 -JOINTS
Pressure
F-POCKET JOtNT SECTION AT
CLIP PUNCH
E -SLIDING SEAM
G-STANDING
High
C- INSIDE GROOVE
SEAM
e- s S L I P
SEAM
AND
S EAMS FOR
Ii
-PITTSBURGH SEAM
Low P RESSURE SYSTEM
Systems
Table I7 contains the construction recommendations for rectangular duct made of aluminum or
steel. The table includes the required reinforcing
and bracing and types of joints and seams used in
high pressure duct systems.
TACK WELDEO
OR RIVETED
Fig. 59 shows the common joint used for rectangular ducts in high pressure systems. The ducts are
constructed with a Pittsburg lock or grooved longitudinal seams (Fig. 58).
Table 15 shows the recommended duct construction for round ducts. The data applies for either
high or low pressure systems. Fig. 60 illustrates the
seams and joints used in round duct systems. The
duct materials for Spira-Pipe
are given in TaOle 16.
1
-
FIG.
i BOLTED
59 -JOINT FOR H IGH P RESSURE S YSTEM
PART 2. AIR DISTRIBUTION
2-62
TABLE 17-RECOMMENDED
CONSTRUCTION
FOR RECTANGULAR SHEET METAL DUCTS
High Pressure Systems
DUCT
MATERIAL
GAGE
Steel
U.S. Gage
Aluminum
B B 5 Gage
Up to 24
22
20
25to 48
20
18
49to 6 0
18
16
DIM
(in.)
61
and Up
18
16
SHEET METAL SCREWS
RECOMMENDED
CONSTRUCTION*
Tranrvarse J o i n t s
Bracing and Reinforcing
n
F l a n g e d a n g l e garketed
loint or butt welded joint
with girth angle, spaced
not more than twelve feet
apart. Angles are 1 I%” x
1 1/l” x %“t.
1 j/z” x 1 l/z” x l/l” girth
angle
reinforcing spaced
38” to 40” apartt.
F l a n g e d a n g l e gasketed
joint or butt welded joint
with girth angle, spaced
not more than twelve feet
apart. Angler are 1 l/2* x
1 ‘/a” x ?&“i.
1 H” x 1 l/i” x 9%” g i r t h
angle reinforcing spaced
38” x 40” aportt.
FIG . G1 -JOINTAND~AM FORSPIRA-PIPE
.
Fittings are normally used to ,join sections of
Spim-Pipe as shown in Fig. 61. Seahng compound is
u s e d to join Spira-Pipe
to fittings.
*All ducts over 18” in either dimension are cross-broken except those to
which rigid board insulation is applied or orea where outlets are installed. Seams are either Pittsburg lock sec~m or longitudinal sec~m.
tAngle
?
“SPIRA-PIPE”SEAM,
are attached to duct by tack welding or rivets on 6” centers.
WEIGHTS OF DUCT MATERIALS
Table 18 gives the weights of various materials
used for duct systems.
COUPLING SLEEVE JOINT
CONTINUOUS WELDED SEAM
CONTINUOUS BUTT WELDED JOINT
GROOVED LONGITUDINAL SEAM
BELL JOINT
Frc.60
-ROUND
DUCT
JOINTSAND~EAMS
CHAPTER 2. AIR DUCT DESIGN
2-63
TABLE II-WEIGHTS OF DUCT MATERIAL
WEIGHT GAGE (THICKNESS)
(in.)
(lb/v ft)
WEIGHT PER SHEET (lb)
36 x 96
48 x 96
48 x 120
GALVANIZED STEEL, U.S. GAGE
.906
1.156
1.406
1.656
2.156
2.656
3.281
~
HOT ROLLED STEEL, U.S. GAGE
,750
1.000
1.250
1.500
2.000
2.500
3.125
5.625
26
24
22
20
18
16
14
10
ga.
ga.
90.
9".
9-z.
ga.
ga.
ga.
1.0179)
I.02391
t.0299)
C.0359)
t.0478)
f.0596)
(.0747)
(.1345)
18.0
24.0
30.0
36.0
48.0
60.0
78.0
135.0
24.0
32.0
40.0
48.0
64.0
80.0
104.0
180.0
ALUMINUM, El 8 5 GAGE ( 3 5 )
.355
.456
.575
.724
.914
1.03
24
22
20
18
16
14
12
ga.
go.
ga.
gel.
9".
ga.
go.
(.020)
t.025)
f.032)
f.040)
f.051)
t.064)
(.071)
6.9
8.6
11.0
13.8
17.4
22.0
24.7
9.2
11.3
14.6
18.4
23.2
29.2
33.0
11.5
14.2
10.2
23.0
29.0
36.6
41.3
STAINLESSSTEEL,U.S.GAGE (302)
.
.66
.79
1.05
1.31
1.58
2.10
2.63
3.28
15.8
la.9
25.2
31.5
37.8
50.4
63.0
70.7
21.1
25.2
33.6
42.0
50.4
61.2
84.0
104.9
26.4
31.6
42.0
52.5
63.0
84.0
105.0
131.2
32.0
40.0
48.0
64.0
72.0
80.0
40.0
50.0
64.0
80.0
90.0
100.0
COPPER, OZjSO FT
1.00
1.25
1.50
2.00
2.25
2.50
16
20
24
32
36
40
OZ.
oz.
oz.
OZ.
or..
oz.
f.0216)
t.027)
f.0323)
(.0432)
t.0486)
(.0540)
24.0
30.0
36.0
48.0
54.0
60.0
2-65
CHAPTER 3. ROOM AIR
This chapter discusses
the distribution of conditioned air alter it has been transmitted to the room.
The discussion includes proper room air distribution, principles of air distribution, and types and
lkation of outlets.
TABLE 19-OCCUPIED ZONE ROOM AIR
VELOCITIES
ROOM AIR
VELOCITY
(fpm)
O-16
REQUIREMENTS NECESSARY FOR GOOD
AIR DISTRIBUTION
25
TEMPERATURE
Recommended standards for room design conditions are listed in Pal-t I, Chupter 2. The air distr:‘>llting system m u s t b e d e s i g n e d t o h o l d t h e
tc
,crature within tolerable limits of the above
recokmendations. In a single space a variation of
2 F at different locations in the occupied zone is
about the maximum that is tolerated without complaints. For a group of rooms located within a
space, a maximum of 3 F between rooms is not unusual. Temperature variations are generally more
objectionable in the heating season than in the
cooling season.
Temperature fluctuations are niore noticeable
than temperature variaions. These fluctuations are
usually a function of the temperature control system.
When they are accompanied by air movements on
the high end of the recommended velocities, they
may result in complaints of drafts.
DISTRIBUTION
1
RECOMMENDED
APPLICATION
REACTION
Complaints
air
Ideal
about
stagnant
design-favorable
all commercial
applications
all commercial
applications
25-50
Probably favorable but 50
fpm is approaching maximum
tolerable velocity for seated
peWXS
65
Unfavorable-light papers
are blown off a desk
75
Upper limit for people moving
aboutslowly-favorable
retail and
dept. store
75-300
Some factory air conditioning
installations-favorable
factory air
conditioning
higher
velocities for
spot cooling
FAIR
t
AIR VELOCITY
Tnble I9 shows room air velocities. It also il-
Ius~~tes occupant reaction to various room air
ve
Lies in the occupied zone.
AIR DIRECTION
Table 19 shows that air motion is desirable and
FAIR
actually necessary. Fig. 62 is a guide to the most
desirable air direction tor a seated person.
I
PRINCIPLES OF AIR DISTRIBUTION
air
The following section describes the principles of
distribution.
BLOW
Blow is the horizontal distance that an air stream
travels on leaving an outlet. This distance is measured from the outlet to a point at which the velocity
of the air stream has reached a definite minimum
value. This velocity is 50 fpm and is measured at
6.5 ft. above the floor.
F1c.62
-DESIRABLE
AIR
DIRECTION
I
2-66
I’.\K.I‘ 2 . .\ I K I~ISTKIBUTION
I%low is 1~roportional to the velocity of the primary
air as it leaves the outlet, and is independent of the
temperature difference bctwecn the supply air and
the room air.
DROP
Drop, or rise, is the vertical distance the air moves
bctwccn the time it leaves the outlet and the time it
reaches the end of its blow.
1NDUCTlON
Induction is the entrainment of room air by the
air ejected from the outlet and is a result of the
velocity of the outlet air. The air coming directly
from the outlet is called primary air. The room air
which is picked up and carried along by the primary
air is crtllcd secondary air. The entire stream, corn- ,
posed of a mixture of primary and secondary air,
is called total air.
Induction is expressed by the momentum
equation:
M, V, + M, V, = (M, + MJ x V.1
where M, = mass of the primary air
M, = mass of the secondary air
V, = velocity of the primary air
V, = velocity of the secondary air
V,) = velocity of the total air
Induction ratio (R) is defined as the ratio of total
air to primary air;
total air
primary + secondary air
R=
=
primary air
primary air
IMPORTANCE OF INDUCTION
Since blow is a function of velocity and since the
rate of decrease of velocity is dependent on the rate
)f induction, the length of blow is dependent on
the amount of induction that occurs. The amount of
induction for an outlet is a direct function of the
perimeter of the primary air stream cross-section.
For two outlets having the same area, the outlet
with the larger perimeter has the greatest induction
and, therefore, the shortest blow. Thus, for a given
air quantity discharged into a room with a given
pressure, the minimum induction and maximum
blow is obtained by a single outlet with a round
cross-section. Conversely, the greatest induction and
the shortest blow occur with a single outlet in the
form of a long narrow slot.
SPREAD
Spread is the angle of divergence of the air stream
after it leaves the outlet. Horizontal spread is di-
vcrgcnce in the horiLonta1 plane and vertical
is divergence in the vertical plane. Spread
incluclccl angle measured in dcgrccs.
Spread is the result ol the momentum law.
is an illustration of the effect of induction on
area and air velocity.
spread
is the
Fig. 63
strcatn
O U T L E T , 1 S O FT
1000
1000
CFM
FPM
2000 CFM
500 FPM VEL
.
F1c.63
-EFFECTOF~NDUCTION
Example 7 - Effect of Induction
Given:
1000 cfm primary air
1000 cfm secondary air
1000 fpm primary air velocity
0 fpm secondary air velocity
Find:
.
The velocity and area of the total air stream when 1000
cfm of primary and 1000 cfm of secondary air are mixed.
Solution:
Area of the initial primary air stream before induction
1000
Ml
=---=~=lSqft
VI
Substituting in the momentum
equation
(1000 x 1000) + (1000 x 0) = (1000 + 1000) v,
v, = 500
Area of the total air stream
=
M, + M,
v,
1000 + 1000
=
500
= 4 sq ft
An outlet discharging air uniformly forward, no
diverging or converging vane setting, results in a
spread of about an 18” to 20” included angle in
both planes. This is equal to a spread of about one
foot in every six feet of blow. Type and shape of
outlet has an influence on this included angle, but
for nearly all outlets it holds to somewhere between
Ijo and 23”.
INFLUENCE OF VANES ON OUTLET PERFORMANCE
Straight Vanes
Outlets with vanes set at a straight angle result
in a spread of approximately 19” in both the horizontal and vertical plane (Fig. 64).
FIG. 65 - SPKEAI)
FIG. 64
c
- S PREAD WI.~H S TRAIGHT
WI’I‘H CONVIXGING
V ANES
VANES
erging Vanes
Outlets with vanes set to direct the discharge air
(Fig. 65j result in approximately the same spread
(19”) as when the vanes are set straight. However,
the resulting blow is approximately 15% longer
than the straight vane setting.
Diverging Vanes
Outlets with vanes set to give an angular spread
to the discharge air have a marked effect on direction
and distance of travel. Vertical vanes with the end
vanes set at a 45” angle, and all other vanes set
at intermediate angles to give a fanning effect, produce an air stream with a horizontal included angle
of approximately 60” (Fig. 66). Under this condition
the blow is reduced about 50y,.‘Outlets with end
vanes set at angles less than 45”, and all other vanes
set at intermediate angles to give a fanning effect,
a blow correspondingly larger than the 45”
hv.
setting, but less than a straight vane setting.
Where diverging vanes are used, the free outlet
area is reduced; therefore, the air quantity is less
than for straight vanes unless the pressure is increased. To miss an obstruction or to direct the air
in a particular direction, all vanes can be set for a
specific angle as illustrated in Fig. 67. Notice that
the spread angle is still approximately 19”.
FIG. 66
- S PREAD WITFI DIVERGING V ANES
INFLUENCE OF DUCT VELOCITY ON OUTLET
PERFORMANCE
An outlet is designed to distribute air that has
been supplied to it with velocity, pressure and
direction, within limits that enable it to completely
perform its function. However, an outlet is not designed to correct unreasonable conditions of flow in
the air supplied to it.
FIG. 67
- S PREAD
WITH
S TRAIGHT V ANES S ET A T
AN A N G L E
2-68
I’.\R’l‘ 2
I
I
LJ
VA = D U C T V E L O C I T Y
Vg= V E L O C I T Y D U E T O P R E S S U R E
DIFFERENCE ACROSS OUTLET
vC = R E S U L T A N T O U T L E T VELOCITY
-It
F1c.68 - OUTLETLOCATEDIN
WITHOUT
VANES
WITH
.\IIi DIS’I‘l<IIIU-I‘ION
Wl~erc an outlet without vanes is located directly
against the side ol‘ a duct, tllc tlirection oE blow 0E
[lie air from the outlet is the vector sum 0E the duct
velocity and the olttlct velocity (Fig. 68). This may
he motlifictl by the 1)cculiarity 0C the duct opening.
Wlicrc an outlet is applictl Lo the ktcc 0C the duct,
the resultant velocity V. can be modified by adjustable V;LIICS behind the outlet. Whether they shbulti
be applied or not depends on the amount of divcrgeticc Irom straight blow that is acceptable.
OEtcn outlets are mounted on short extension
collars away from the lace of the duct. Whenever the
duct vciocity exceeds the outlet discharge velocity,
vanes should bc used where the collar joins the duct.
Results are indicated in Fig. 69.
DUCT
.
IMPORTANCE OF CORRECT BLOW
-
VANES
F1c.69 -COLLARFOR~UTLETS
SUPPLY AIR WARMER
THAN ROOM AIR
.
\
Normally it is not necessary to blow the entire
length or width of a room. A good rule oE thumb
to follow is to blow 3/4 oE the distance to the opposite
wall. Exceptions occur, however, when there are
local sources of heat at the end of the room opposite
the outlet. These sources can be equipment heat
and open doors. Under these circumstances, overblow may be required and caution must be gxercised
co prevent draft conditions.
SUPPLY TEMPERATURE DIFFERENTIAL
The allowable supply temperature difference that
can be tolerated between the room and the supply
air depends to a great extent on (1) outlet induction
ratio, (2) obstructions in the path of the primary air,
and (3) the ceiling height. Fig. 70 indicates the effect
of changing the supply air temperature from warm
to cold.
Since induction depends on the outlet velocity,
there is a supply temperature differential which
must be specified to give satisfactory results.
TOTAL ROOM AIR MOVEMENT
SUPPLY AIR
EQUALS
ROOM AIR
TEMP
SUPPLY AIR COOLER
THAN ROOM AIR
The object of room air distribution is to provide
satisfactory room air motion within the occupied
zone, and is accomplished by relating the outlet
characteristics and performance to the room air motion as follows:
1. Total air in circulation
= outlet cfm
x
induction ratio.
2. Average room velocity
FIG. 70 -AIR STREAM PATTERNS
TEAWERATURE
FOR
DIFFERENTIALS
VARIOUS
1.4 X total cfm in circulation
= area of wall opposite outlet(s)
3. K=
average room velocity
1.4 X induction ratio
outlet cfm
= clear area of wall opposite outlet(s)
where K is the room circulation factor expressed in primary air cfm/sq ft of wall opposite the outlet.
The multiplier 1.4 allows for the blocking caused
by the air stream. Note that the clear wall area is
indicated in the equation and all obstruction must
be deducted. See Note 8, Table 21.
Table 19 indicates that the average room air movement should be kept between 15 and 50 fpm for
most applications. Tests have been performed on
outlets at various outlet velocities to determine perCc ante characteristics. The results of such tests
01. . jpecifi’c series of wall outlets (Fig. 93) are shown
in the rating tables at the end of this chapter. This
rating data can be successfully used for outlets having the nominal dimensions and free area indicated
in Table 21. An example illustrating outlet selection
accompanies the table. The K factor as indicated
in Item 3 is shown at the bottom of the rating table
as maximum and minimum cfm/sq ft of outlet
wall area.
TYPES OF OUTLETS
PERFORATED
GRILLE
This grille has a small vane ratio (usually from
0.05 to 0.20) and, therefore, has little directional
effect. Consequently, it is used principally as an
exhaust or return grille but seldom as a supply
grille. When a manual shut-off damper backs up this
gT;‘b) it becomes a register.
Fl);cti BAR GRILLE
The fixed bar grille is used satisfactorily in locations where flow direction is not critical or can be
predetermined. A vane ratio of one or more is desirable. To obstruct the line of sight into the duct
interior, closely spaced vanes are preferred.
ADJUSTABLE BAR GRILLE
This grille is the most desirable for side wall
location. Since it is available with. both horizontal
and vertical adjustable bars, minor air motion problems can be quickly corrected by adjusting the vanes.
SLOTTED OUTLET
This outlet may have multiple slots widely spaced,
resulting in about 10% free area. Performance is
about. the same as for a bar grille of the same cfm
and static pressure, but the blow is shorter because
of greater induction at the outlet face.
Another design to effect early completion of induction is the long single, or double, horizontal slot.
It is particularly advantageous where low ceiling
heights exist and outlet height is limited, or where
objections to the appearance of grilles are raised.
EJECTOR OUTLET
The ejector outlet operates at a high pressure to
obtain a high induction ratio and is primarily used
for industrial work and spot cooling. When applied
to spot cooling, a high degree of ejector flexibility
is desired.
INTERNAL INDUCTION OUTLET
Where a sufficiently high air pressure is used,
room air is induced thru auxiliary openings into the
outlet. Here it is mixed with primary air, and discharged into the room at a lower temperature differential than the primary stream. Induction progresses in two steps, one in the outlet casing and the
other after the air leaves the outlet.
CEILING OUTLETS
Pan Outlet
This simple design of ceiling distribution makes
use of a duct collar with a pan under it. Air passes
from the plenum thru the duct collar and splashes
against the pan. The pan should be of sufficient
diameter to hide the duct opening from sight, and
also should be adjustable in distance from the ceiling. Pans may be perforated to permit part of the
air to diffuse downward. Advantages of the pan outlet are low cost and ability to hide the air opening.
Disadvantages are lack of uniform air direction because of poor approach conditions and the tendency
to streak ceilings.
Ceiling Diffuser
These outlets are improvements over the pan
type. They hasten induction somewhat by supplying
air in multiple layers. Approach conditions must be
good to secure even distribution. Frequently they
are combined with lighting fixtures, and are available with an internal induction feature. See Fig. 71.
Perforated Ceilings and Panels
Various types of perforated ceilings for the introduction of conditioned air for comfort and industrial systems are available. The principal feature
of this method of handling air is that a greater
volume of air per square foot of floor area can be
introduced at a lower temperature, with a minimum
of movement in the occupied zone and with less
I
I’\ll
2-70
FIG.
71 - INTERNAL~NDWXION
danger of draft. Since discharge velocity is low,
induction is low. Therefore, care must be taken to
provide adequate room air motion in excess of 15
fpm.
Duct designed for a perforated ceiling is the same
as duct designed for a standard ceiling. To obtain
adequate supply to all areas, the same care necessary
for conventional systems must be taken in laying
out ducts for the perforated ceiling. The ceiling
panels should not be depended upon to obtain
proper air distribution, since they cannot convey air
to areas not otherwise properly supplied. Perforated
panels do assist in “spreading out” the air supply
and, therefore, comparatively large temperature differentials may be used, even with low ceiling heights.
APPLICATION OF CEILING DIFFUSERS
Installations using ceiling diffusers normally result in fewer complaints of drafts than those using
side wall terminals. To eliminate or minimize these
complaints, the following recommendations should
be considered when applying ceiling diffusers.
BLOW
Select ceiling diffusers for a conservative blow,
generally not over 75% of the tabulated value. Overblow may cause problems on many installations;
under-blow seldom does.
.
CULISG
1‘ ‘2. \II<
DISTKIBUTION
DI~FUSEK
PRESSURE DROPS
Most rating tables express the pressure drop thru
the outlet only and do not include the pressure
drop necessary to force the air out of the duct thru
the collar and outlet and into the room. Therefore,
it is recommended that rated pressure drops be carefully investigated and the proper safety factor applied Tvhen necessary.
DIFFUSER APPROACH
,111 important criterion for good diffuser performance is the proper approach condition. This means
either a coilar of at lcast -1 times the duct diameter,
or good turning vmcs. If vanes are used, they must
be placed perpendicular to the air flow at the upper
end of the collar 2nd spaced approximately 2 in.
apart.
OBSTRUCTIONS
\\%el-e obstructions to the flow of air from the
clilf~laer occur, blank off a small portion of the diffuser at the point at whicll the obstruction is located.
Clip-on baffles arc‘ us11:111y provided for this purpose.
OUTLET NOISE LIMITATIONS
One important criterion affecting the choice of an
outlet is its sound level. Table 20 shows recommen&d outlet velocities that result in acceptable
jountl levels for various types of applications.
(:l-l,\l’~l‘l~K
3.
I1001Ll
.\II<
TABLE 20 -RECOMMENDED OUTLET
VELOCITIES
TERMINAL VELOCITY
IFPM)
APPLICATION
Broadcast studios
Residences
Apartments
Churches
Hotel bedrooms
Legitimate
theaters
Private officer, acoustically
Private offices, not treated
Motion picture theaters
General offices
Dept. storer, upper f!oors
Dept. stores, main floor
OUTLET
300-500
500-750
500-750
500-750
500-750
500-750
500-750
500-800
1000
treated
1000-l 250
1500
2000
LOCATIONS
Interior architecture, building construction rend
streaking possibilities necessarily influence the
1‘. , .ut and location of the outlet. However desirable
(!
FIG. 72 - DOWNDRAFT
W
2-71
I>IS’I‘RIIIIJ’I’IOU
I N D O W
FROM
COLD
it may I)e t o loc;lte 211 outlet in 2 given .sl)ot, these
items may prevent such locxtion.
A f t e r 211 the forcgoin,g l i m i t a t i o n s have been
successfully tle;tlt with, the 2ir distribution principles whic,h relate to flow, drop, c;ipicity xntl r o o m
air circul;ltion create further limitations in tlesigning an xcccpt;~l)le air distribution system. These arc
t;lbulatetl i n t h e lating t:kt)ies a t the end o f t h e
chapter.
L o c a l loads due t o people concentration, equipment heat, outside walls ant1 window locations frequently modify the choice of outlet location. The
downdraft from a cold wall or a glass window (Pig.
72) can reach velocities of over 200 fpm, causing
discomfort to occupants. Unless this downdraft is
overcome, complaints of cold feet result. In northern
climates this is accomplished by supplementary radiation, or by an outlet located under a window as
illustrated in Fig. 73.
FIG. 73 -DISCHARGE A IR OFFSETTING WINDOW
DOWNDRAFT
I’:\ li.1‘ 2.
2-72,
Another item to consider wllcn choosing an outlet
location is the radiant ellcct from cold or warm
surfaces.
During the heating season an outlet tlischarging warm air under a cold window raises the
surface temperature and rcduccs
the Lceling of discomfort.
The following describe four typical applications
of specific outlet types.
CEILING DIFFUSERS
Ceiling diffusers may be applied to exposed duct,
furred duct, or duct concealed in a ceiling. Although
wall outlets are installed on exposed and furred
duct, they are seldom applied to blow directly downward unless complete mixing is accomplished before
the air reaches the occupied zone.
WALL OUTLETS
A high location for wall outlets is preferred where
a ceiling is free from obstructions. Where beams
are encountered, move the outlet down so that the
air stream is horizontal and free from obstruction.
If this is not done and if vanes are used to direct the
air stream downward, the air enters the occupied
zone at an angle and strikes the occupants too
quickly. This is shown in Fig. 74.
Wall outlets located near the floor (Fig. 75) are
suitable for heating but not for cooling, unless the
air is directed upward at a steep angle. The angle
must be such that either the air does not strike
occupants directly or the secondary induced stream
does not cause an objectionable draft.
WINDOW OUTLETS
Where single glass is used, window outlets are
r
AIR
STREAM
2
BLOW
HAS
A
THAN
.\IR I)IS’I‘KlI~UI‘ION
prcl’erretl to cithcr wall or ceiling distribution to
oll‘set the pronoiinccd tlowntlraTt during Lhc winter.
7‘11~ air slwuIt1 be directed with vanes at an angle
ol 15” or 20’ from the vertical into tile room.
FLOOR OUTLETS
Where people are seated as in :I theater, floor
outlet distribution is not permissible. Whcrc people
are walking about, it is possible to introduce air at
the floor level; for cxainplc, in stores where air is
directed horizontally thru a slot under a counter.
In this application, liowevcr, a very low temperaturc tliffercntial of not more than 5 or 6 degrees must
be used. Maintaining this maximum is usually uncconornical b e c a u s e o f the large air volume rcquired. However, if air is directed upward behind
the counter and diffused at an elevation above the
occupied zone, the temperature differential may be
increased approximately 5 times. Another disadvantage is that floor outlets become dirt collectors.
SPECIFIC
APPLICATIONS
If the principles described in the previous paragraphs are properly applied, problems after installation will be at a minimum. Basically, the higher the
ceiling the fewer the number of problems encountered, and consequently liberties may be taken at
little or no risk when designing the system. However, wiih ceiling heights of approximately 12 feet
or less, greater care must be exercised to minimize
problems.
Experience has shown that ceiling diffusers are
easier to apply than side wall outlets, and are preferred when air quantities approach 2 cfm/sq ft
of floor area.
The following general remarks about specific ap-
OBSTRUCTION
LONGER
AIR
EFFECTIVE
STREAM
1
,.
F L O O R
.
j ‘-
:, :
.’
FIG. 74 -WALL OUTLET
IN
ROOM
O BSTRUCTION
W ITH CEILING
FIG. 75 - W ALL OUTLET NEAR FLOOR
l
(:H:\I’TEK 3.
ROOM
2-73
,\IK I~IS’I‘RII1U’I‘I<)N
plicatiotls arc the result of thousands of installations
and are olfercd ;is a guide for better air distribution.
Apartiileiits, hotels and olfice I)uildings are discussed
ii1 relation to specilic location of s o u r c e s of air
supply con~mon to these
types of buildings. 12anks,
restaurants, and department and specialty stores are
discussed in more general terms, although the common sources of locations of outlets previously dis-
5
.
Ilctiirii grille:
ixturn xir tllrri tile c.orritlor
arid return ducts arc not used,
CVlierc
u s e r e l i e f grilles
o r t o untlcrcut
i
s
I)crltlissil)le
it is nccc~ssary
to
olliw doors.
In apartments and hotels, cotlcs must 1)~’ checked
bcforc using the corridor, as a return plcnuni. Even
if codes permit, it is not good engineering practice
to me the corridor as a return plenum.
cussctl c a n be a p p l i e d .
APARTMENTS, HOTELS AND OFFICE BUILDINGS
1. Corridor Supply - No direct radiation (Fig. 76):
A dvur1 loge - Low cost.
Lhadvantage - Very poor in winter. Downdraft
under the window accentuated by the outlet blow.
Precazllion - Blow must be not more than 75y0
of the room length.
2 Corridor Supply - Direct radiation under win.ows (Fig. 77):
Advnntuge - Offset of downdraft under window
in winter when the heat is on.
Disadvantage - Slight downdraft still occurs during intermediate season or whenever radiation is
shut off during cool weather.
Precaution - Do not blow more than 75% of
.
room length.
3. Duct above window blowing toward corridor
(Fig. 78):
Advantage - Somewhat better than corridor distribution but does not prevent winter downdraft
unless supplemented by direct radiation.
Disadvantage - Nearly as expensive (considering
building alterations) as window outlet which results in better air distribution.
4. Window outlet (Fig. 79):
Advantage - Eliminates winter downdraft - by
“ar the best method of distribution.
3isndvantage - May not be economical for multiple windows.
ELEVATION
FIG.
77 - CORRIDOR
AIR S UPPLY
WITH
.
I
FIG.
78 - DC’CT
ABOVE W INDOW , B LOWING TO~VARD
CO R R I D O R
_
r-l
E L E V A T I O N
ELEVATION
7 6 - CORRKJOR
AIR S UPPLY
1
ELEVATION
-
I
DIRECT
RADIATION
Flc. 59 -
WIMOW
OUTLET
,
2-74
PART
BANKS (FIG. 80)
Often the center bank space has a high ceiling
with an electrical load. In this case, use of side wall
outlets part of the way up the wall may result in
segregating some of the ceiling load and keeping it
out of the occupied zone, thus permitting some reduction in cooling load. This location of outlets part
way up the wall is used with ceiling heights in
excess of 20 ft.
DEPARTMENT STORES (FIG. 81)
There is nothing critical about air distribution in
department stores if ordinary precautions are observed, provided the ceiling is high enough. Care
should be taken in conditioning a mezzanine since
the outlet is likely to overblow and not cool its occupants. Longitudinal distribution is preferred. Base-
2
.
.\lR DISTRIHUTION
merits may give trouble clue to low ceilings and pipe
obstructions. Main floors usually rquire more air
near doors.
RESTAURANTS (FIG. 82)
Great care must be taken in locating outlets with
respect to exhaust hoods or kitchen pass-thru
windows. Usually the air velocities over such openings
are low and excessive disturbance due to direct blow
or induction from outlets may pull air out of them
into the conditioned space.
STORES
1. Outlets at rear blowing toward door (Fig. 83):
Requirements - Unobstructed ceiling.
Disadvanta,q - May result in high room circulation factor K.
Precaution - Blow m’ust be sized for the entire
KITCHEN
WRONG
.
ELEVATION
FIG. 80
t
- A IR DISTRIBUTION
WITH
HIGH CEILING
-RIGHT
I
I
PLAN
WRONG
FIG . 82 - RESTAURANT A IR DISTRIBUTION
RIGHT
-
ELEVATION
FIG.
81
- MEZZANINE A IR DISTRIBUTION
PLAN
FIG .
83 - A IR DISTRIBUTION
FROM
REAR
OF
S TORE
.
(:H.\I’TEK
2.
3.
4.
-_
3. R O O M
2-75
.\IR DISTRIRUTION
length of the room; otherwise a hot zone may
occur due to infiltration at the doorway. Care
must be taken to avoid downdrafts near walls.
Outlets over door blowing toward rear (Fig. 84):
Requirements - Unobstructed ceiling.
Disadvnnlnge - May result in high room circulation.
Precaution - Excessive infiltration may occur
due to induction from doorway.
Outlets blowing from each end toward center
(Fig. 85):
Advantage - Moderate room circulation.
Pwcazdion - There may be a downdraft in the
center. Outlets should be sized for blow not
greater than 40 percent of the total length of the
room.
Center outlets blowing toward each end (Fig. 86):
4duantage - Moderate room circulation.
Juct along side wall blowing across the store
Adwmtnge - ISest air distribution.
I1is/lcl71(ln&rge - High cost.
THEATERS
1. Ejector system for sma!l theaters, no baIcony
(Fig. 89):
Re~Iuirements - Unobstructed ceiling 2nd ability
to locate outlets in the rear wall.
Acl7xiritage - Low cost.
Precaution - Possibility of clcatl
spots at front
back of theater. Use mushrooms for return
air under seats if excavated. In northern climates
direct radiation may be advisable along the sides.
xltl
-
(Fig. ZT):
Advantage - Moderate room circulation.
Precaution - Overblow may cause downdrafts on
the opposite wall.
6. Ceiling diffusers (Fig. 88):
Requil-ements - Necessary where ceiling is badly
cut u p .
FIG . 86 - A IR DISTRIBUTION
FIG. 87
- A IR DISTRIBUTION
CENTER
FROM
S IDEWALL OUTLETS
FROM
CEILING DIFFUSERS
OF
S TORE
PLAN
PLAN
FIG. 84
FROM
FROM
OVER
THE
- A IR DISTRIBUTION
DOOR
-
PLAN
FIG .
85 - A IR DISTRIBUTION
OF
STORE
FROM
E ACH E ND
FIG.
88 - A IR DISTRIBUTION
2-76
I’ART
2.
.\IR
DISTRII3UTION
L
d &jnyf tnge - ConiplcLc coverage, no tlcatl spots.
Ilistrtl7urrrtfrge - I-1 igher first cost.
~~-ecll~l/io~~ - i\ir must not stnikc obstructions
wit11 a velocity force that causes tlellcction and
drafts in the occupiccl zone. Temperature tlillerentials must be limited in regions of low ceiling
heights. USC low outlet vclocitics.
ELEVATION
FIG. 89 -AIR DISTRIUUTION
FOR
SMALL THEA?XES
2. Ejector system for large theaters with balcony
(Fig. 90):
Requiwrnents
- Unobstructed ceiling.
- Low cost.
Precaution - Balcony and orchestra should have
separate returns. Preferred location, under seats;
acceptable location, along sides or rear of theater.
Return at front of theater generally not acceptable. Outlets under balcony should be sized for
distribution and blow to cover only the area
directly beneath the balcony. Orchestra area
under balcony should be conditioned by the
balcony system. Allow additional outlets in rear
for standees when necessary.
3. Overhead system (Fig. 91):
Requirements - Necessary when ceiling is obstructed.
A
dunn tage
RETURN
Velocities thru return grilles depend on (I) the
static pressure loss allowed and (2) the effect on
occupants or materials in the room.
In determining the pressure loss, computations
should be based on the free velocity thru the grille,
not on the fact velocity, since the orifice coefficient
may approach 0.7.
In general the following velocities may bc used:
GRILLE LOCATION
FPM OVER
GROSS AREA
Commercial
Above occupied zone
800 and above
Within occupied zone not nea; seats
Within occupied zone near seats
Door or wall louvers
Undercutting of doors
Industrial
Residential
600-800
400-600
poo-1000
600x
800 and ahove
400
*
*Thru
L-
GRILLES
undercut
area
LOCATION
BALCONY
ELEVATION
F1c.90 -AIRDISTRIBUTION
FOR
LARGE THEATRES
WITHBALCONY
THEATER
ELEVATION
FIG. 9 1 - OVERHEAD
AIR DISTRIBUTION
Even though relatively high velocities are used
thru the face of the return grille, the approach
velocity drops markedly just a few inches in front
of the grille. This means that the location of a return grille is much less critical than a supply grille.
Also a relatively large air quantity can be handled
thru a return grille without causing drafts. General
drift toward the return grille must be within acceptable limits of less than 50 fpm; otherwise complaints resulting from drafts may result. Fig. 92
indicates the fall-off in velocity as distance from the
return grille is increased. It also illustrates the approximate velocities at various distances from the
grille, returning 500 cfm at a face velocity of 500
fpm.
Ceiling Return
These returns are not normally recommended.
Difficulty may be expected when the room circulation due to low induction is insufficient to cause
warm air to flow to the floor in winter. Also, a
poorly located ceiling return is likely to bypass the
cold air in summer or warm air in winter before it
has time to accomplish its work.
,
(:H:\I”1‘EK
3.
l<OOM
.\IR
2-77
I~IS’I‘KIIIIJ~I‘ION
OUTLET SELECTION
The following example describes a method of
selecting a wall outlet using the rating tables on
page 78.
Example 2 - Wall Outlet Selection
Given:
Small store
Dimensions - 32 ft x 23 ft x 16 ft
Ceiling - flat
Load - equally distributed
Air quantity - 2000 cfm
Temp diff - 25 F
Find:
Number of outlets
Size of outlets
Location
PLAN
FIG. 92-VELocrrY FALL-OFF
FROM GRILLE
Wall
PER
DISTANCE
Return
A wall return near the fldor is the best location.
Wall returns near the ceiling are almost as undesirable as ceiling returns. Differences due to poor
mixing in the winter are counteracted by a low
return since the cool floor air is withdrawn first
and is replaced by the warmer upper air strata.
Floor Return
These should be avoided wherever possible because they are a catch-all for dirt and impose a
severe strain both on the filter and cooling coils.
In7 -never floor returns are used, a low velocity sett: J chamber should be incorporated,
Solution:
First, find the required blow in feet and the wall outlet
area (air movement K factor). The minimum blow is 75%
of the room width for the given condition of equally distributed heat load. Therefore, the minimum blow necessary
is s/d X 23 = 17.3 ft. The maximum blow is the width of
the room. The outlet wall area K factor is- equal to the
cfm supplied divided by the outlet wall area:
2000
i
= 3.9 primary air cfm/sq ft wall area
32 X 16
Enter Table 21 and select one or more outlets to give a
blow of at least 17.3 ft. Air movement must be such that
the value of K equals 3.9 primary air cfm/sq ft and that this
value falls within the maximum and minimum values which
are shown at the bottom of the rating tables. The tables indicate that, to best satisfy conditions, four outlets, nominal
size 24 in. x 6 in., are to be used. By interpolating, it is
found that the four 24 in. x 6 in. outlets at 500 cfm have a
range in blow of 17.5 to 34 ft. By adjusting the vanes the
proper blow can be obtained. Also the velocity of the outlet is found to be about 775 fpm. This is well within the
recommended maximum velocity of 1500 fpm, Table 20.
The minimum ceiling height from the table is just over 9
ft. This is less than the height of the room: therefore, the
outlet selection is satisfactory. The top of the outlets should
be installed at least 12 in. from the ceiling, (Note 8,
Table 21).
FIG. 93-bVAu OUTLET ON WHICH RATINGS ARE BASED
TABLE 21-WALL
OUTLET RATINGS, FOR COOLING ONLY
For Flat Ceilings
OUTLET
2 5 0 FPM
VELOCITY
STATIC PRESSURE WITH
M E T E R I F4G, P L A T E
Nom. Size
of Outlet
VWle
(and Free
Setting
Area)
500
FPM
Str B = ,013, 22%” = ,015
45O = ,019
750
FPM
Str B = ,024, 22%’ = ,028
45" = ,035
S t r B = . O l , 2 2 ’ / 2 ’ = , 0 1 5 1S t r B = , 0 2 4 , 2 2 ’ / i ” = . 0 4 3 1 S t r B = . O $ l ,
45'.= .02B
4 5 " z-z
I
Air
Air
rem
Difl
Dif f(
Tern
- bJan
.> ouan15
20
15
20
my
_t i t y
Mi
(cfml
Mi
Clg It
(cfm)
- Clg
-
t-
22%’ = .OB2
a.0
Straight
22x0
450
30
-z
2.5
1.8
6.5
6.5
6.0
7.0
6.5
6.0
-
-E5.1
3.5
7.5
6.5
6.5
7s
7.0
6.5
7.0
7.0
10 x 4
(21.7)
Straight
22'/2O
450
37
3.5
2.5
I.8
6.5
6.5
6.0
7.0
6.5
6.0
-
7.4
5.5
3.7
7.5
6.5
6.5
7.5
7.0
6.5
7.0
7.0
Straight
221/2O
45O
44
3.5
2.5
1.8
7.5
7.0
6.5
7.5
7.0
6.5
7.5
7.0
Straight
22x0
45O
61
3.7
2.7
2.0
7.0
6.5
6.5
7.0
6.5
6.5
-
7.5
5.5
3.9
16 x 4
(35.9)
6.5
6.5
6.0
7.0
6.5
6.0
-
7.9
6.0
4.0
7.5
7.0
6.5
7.5
7.0
6.5
7.5
7.0
20 x 4
(45.5)
Straight
22x0
45O
77
4.0
3.0
2.0
7.0
6.5
6.0
a.0
6.0
4.0
7.5
7.0
6.5
a.0
7.0
6.5
7.5
7.0
24 x 4
(55.0)
Straight
22%"
45O
93
4.1
3.1
2.0
7.0
6.5
6.0
7.0
6.5
6.5
7.0
7.0
6.5
a.0
6.0
4.0
7.5
7.0
6.5
30 x 4
(68.3)
Straight
22Yi0
45O
116
4.2
3.1
2.1
7.0
6.5
6.0
7.0
7.0
6.5
-
a.0
6.0
4.0
36 x 4
(83.5)
Straight
22s0
45O
140
4.4
3.3
2.2
7.0
6.5
6.0
7.5
7.0
6.5
8x6
(26.5)
Straight
22'/2O
450
52
0
3.8
2.5
10 x 6
(34.0)
Straight
22%"
45O
66
12 x 6
(41.6)
Straight
22'h0
45O
16 x 6
(56.6)
ran
Diff (F)
15
5rpc
M in Clg Ht
FPM
s t r B = , 0 5 1 , 2 2 ' / 2 0= . 0 6 1
45O = .OB
Str B = .175, 22%” = .I9
45" = .27
-
r
8x4
(16.9)
Air
. Clum - I
tity
IIcfm)
-
3low
vu
r
Tern0 D i f f z
(F)
25
Ht
59
lo.0
7.5
5.0
7s
7.0
6.5
a.0
7.5
6.5
8.5
7.5
7.0
a9
17.0
13.0
9.0
8.5 9.0
7.57.5
6.5 7.0
9.0
a.0
7.0
75
10.5
8.0
5.4
7.5
7.0
6.5
a.0 a.5
7.5
7.5
6.5
7.0
112
18.0
13.0
9.0
8.5 9.0
7.58.0
6.5 7.0
x7
a.0
7.0
91
II.0
8.1
5.5
a.0
7.0
6.5
8.0
7.5
7.0
a.5
7.5
7.0
136
I a.0
13.0
9.0
9.5
a.5
7.0
122
11.0
8.1
5.5
iz
7.0
6.5
8.0
7.5
7.0
a.5
7.5
7.0
183
19.0
14.0
10.0
9.5
a.5
7.5
154
1I . 5
8.5
6.0
G
7.5
6.5
8.0
7.5
7.0
8.5
a.0
7.0
231
20.0
15.0
10.0
9.5
a.5
7.5
s.0
7.0
6.5
8.0
7.5 185
7.0
11.5
a.5
6.0
ii
7.5
6.5
8.0
7.5
7.0
8.5
a.0
7.0
276
20.0
15.0
10.6
10.0
a.5
7.5
7.5
7.0
6.5
iii
7.5
6.5
8.0
7.5 233
7.0
12.0
9.0
6.0
8.0
7.5
6.5
a.0
7.5
7.0
8.5
a.0
7.0
349
21.0
16.0
II.0
Gi
a.5
7.5
8.0
6.0
4.0
7.5
7.0
6.5
7.5
6.5
7.5 279
7.0
a.0
12.0
9.0
6.0
7.5
6.5
420
21.0
16.0
II.0
10.0
8.5
7.5
G7.0
4.8
8.0
7.0
6.5
-
G
7.5
7.0
a.5
8.0
7.0
lb3
KY
10.0
6.0
s.f,
7.5
7.0
155
24.0
I a.0
12.0
10.5
9.5
a.0
5.5
4.1
2.8
7.5 r.s
7.0
7.0
6.0 6.5
7.5
8.0
7.0
7.5
6.5
7.0
10.0
7.5
5.0
a.0
7.5
7
.0
-
a.r,
8.0
7.0
9.0
8.5
7.5
15.0
11.0
7.0
G
131
196
27.0 I
20.0
14.0
II.5
10.0
8.0
7.5
7.0
6.5
-
8.0
7.5
7.0
-
11.0
a.1
5.5
8.0
7.5
7.0
9.0
a.5
238
7.5
15.0
II.0
7.0
G
a.0
9.5
80
6.0
4.5
3.0
28.0 1
21.0
? 4.0
11.5
10.0
- a-.-0L
Straight
221/2O
45O
8.0
7.5
7.0
12.0
9.0
6.0
a.5
a.0
7.0
9.5
8.5 214
7.5
16.0
12.0
a.0
321
7.5
30.0 1
22.0
15.0
12.5
10.5
a.5
20 x 6
(71.5)
Straight
22Yi0
45O
8.5
7.5
7.0
12.0
9.0
6.6
9.0
a.0
7.0
-
9.5
0.0
9.0
7.5
17.0
13.0
9.0
c5
8.5
7.5
1
135
8.0
7.0
6.5
8.0
7.5
7.0
-
9.0
107
6.2
4.7
3.2
403
32.0 1
24.0
16.0
iiT
11.0
9.0
24 x 6
(86.5)
Straight
22%"
450
8.5
8.0
7.0
-
13.0
10.0
6.5
0.0
9.0 324
i2
a.0
a.5
7.5
I 0.511.0
9.010.0
a.0 a.5
466
7.5
Ia . 0
13.0
9.0
33.0 1
25.0 1
17.0
13.0
11.0
9.1
30 x 6
(109.0)
Straight
22%"
45O
203
8.5
7.5
7.0
a.5
a.0
0.0
9.0
7.5
0.5
9.0 406
19.0
14.0
10.0
0.0
9.0
7.5
1 1.0 II.5
9.510.0
a.0 8.5
609
7.0
13.0
IO.0
6.5
9.0
8.0
7.0
9.0
a.0
7
.5
-
95
162
8.0
7.5
7.0
-
34.0 I
25.0 1
17.0
13.5
11.5
9.0
36 x 6
(131.3)
Straight
22%"
45O
245
8.5
7.5
7.0
-
9.0
8.0
7.5
-
13.0
IO.0
6.5
-
9.5
a.5
7.5
-
iz
9.0
0.5
9.5 490
19.0
14.0
10.0
-
iz 1 1.0 12.0
9.0
9.510.0
8 . 01 a . 0 a . 5
-
735
35.0 1
26.0 1
18.0
-
14.0
11.5
9.5
-
12 x 4
(24.6)
'.
375
Str B = .Ol 22’/1” = .Ol
4;o = .Ol
STATIC PRESSURE
STANDARD OUTLET
-
-
6.6
5.0
3.2
7.5
7.0 139
6.5
9.0
8.0 202
7.5
K
a.0
7.0
a.0
7.0
a.5
7.5
a.5
a.0
-
a.0
a.0
a.0
a.0
159
269
a.0
a.0
a.0
-
a.0
7.0
-
a.0
7.0
9.5
a.5
FACTOR
M a x C f m / S qF t
O u t l e tW a l l Area
29.0
19.0
14.0
M i n C f m / S qF t
O u t l e tW a l l Area
a.7
5.7
4.2
1
c
-
9.6
2.9
'
(:M/\I’rI‘EK
3.
ROOM
.\II<
I>IS’l‘IllIllJ’l‘ION
TABLE
2-79
2LWALL OUTLET RATINGS, FOR COOLING ONLY (Cont.)
For Fiat Ceilings
OUTLET
VELOCITY
2000 FPM
STATIC
PRESSURE
STANDARD
OUTLET
S t r 8 = .375, 22%” = .42
45’ = .565
STATIC PRESSURE WITt
METERING
PLATE
Nom. Sire
of Outlet
(and Free
Area)
_,
VatI*
Setting
8 x 4
(16.9)
Straight
22%”
45O
10 x 4
(21.7)
Straight
221/2O
45O
12 x 4
(24.6)
Straight
22%O
45O
Air
-C hantity
I :cfm)
118
237
-
150
224
299
181
10.0
9.0
7.5
272
362
x 4
.5.9)
Straight
22vi”
45O
244
10.5
9.5
7.5
366
488
65
49
33
20 x 4
(45.5)
Straight
22x0
4s”
308
462
616
67
50
34
24 x 4
(55.0)
Straight
22x0
45O
556
740
30 x 4
(68.3)
Straight
22s”
45O
36 x 4
(83.5)
Straight
22’/20
4s”
558
8 x 6
(26.5)
Straight
22x0
45O
206
36
27
18
9.5
9.0
7.5
10 x 6
(34.0)
Straight
22%”
45”
262
40
30
20
12 x 6
(41.6)
Straight
22%”
45O
318
41
16 x 6
‘26.6)
Straight
22%”
45O
428
20 x 6
(71.5)
Straight
22x0
45O
24 x 6
(86.5)
Straight
22x9
45”
648
30 x 6
(109.0)
Straight
22’/2O
45O
812
36 x 6
(131.3)
Straight
22%”
45O
980
Cfm/Sq
Ft
Outlet Will krea
Min Cfm/Sq F t
Outlet Wall Area
12.0
10.0
7.5
11.5
9.5
7.5
12.5
10.0
12.5
10.5
8.0
68
51
34
11.0 12.0
9.0 10.0
7.5
7.5
11.0 12.0
9.5 10.0
7.5
8.0
11.5 12.0
9.5 10.0
7.5
8.0
-
932
70
53
35
11.5
9.5
7.5
-
12.5
10.0
8.0
13.5
11.0
8.5
71
53
36
11.5 12.5
9.5 10.5
7.5 8.0
12.5 14.0
10.5 11.5
9.0
8.5
! 7.0
! 7.5
370
30
22
15
9.5
8.5
7.0
10.0
9.0
7.5
10.5
9.5
7.5
466
30
22
15’
9.5
8.5
7.0
10.0
9.0
7.5
10.5
9.5
7.5
31
9.5 10.0
23
8.5
9.0
16 ! 7.0 ! 7.5
11.0
9.5
8.0
840
1116
11.0
9.5
8.0
12.0
10.0
8.0
310
412
11.0
9.5
8.0
12.0
10.0
8.5
13.0
11.0
9.0
392
-
524
t
31 10.0 10.5
11.5
12.5
13.5
11.0
9.0
476
636
44 12.0 13.0
3 3 I1 0 . 0 I 1 1 . 0
22 ! 8.0 ! 9.0
14.0
11.5
9.0
642
856
47 12.5 13.5
3 5 I1 0 . 5 I 1 1 . 5
14.5
12.0
9.5
t
21
538
! 8.0
! 8.5
H
36
24
10.5
8.5
806
15.5
12.5
9.5
076
972
296
5 0 1 3 . 0 1 4 . 5 15.5
38 11.0 12.cI 12.5
25
9.0
9.5 I 10.0
1218
1624
5 1 1 3 . 5 1 5 . c I 16.0
38 11.0 12.cI 13.0
26
9.0
9.5 110.0
A
1470
1960
7.2
2.2
11.5
9.5
698
K
?
6 0 10.5
45
9.0
30 7.0
62 IO.5
47
9.0
31
7.5
- -
82
62
41
92 14.0
6 9 11.5
46
9.0
-
4.8
has vertical
lowres straight f o r w a r d i n the
center, with uniformly incrsosing
ongulor d e f l e c t i o n lo o moximum ot each end. The 4.5’ div e r g e n c e signifies o n ongulor
deflection at each
end
of the
outlet of 45’. and simitorly
for
22%’ divergence.
5. Velocity
1.4
6.
on effective
Static Pressure is that pressure
required to produce the indi-
The
Minimum
Ceiling
Height
(table) is the minimum ceiling
height which will give proper
operation of the outlet for rho
17.0
13.5
10.0
18.0
14.5
108 16.0
81 13.0
5 4 IO.0
-
17.5
14.0
10.5
19.0
15.0
III
83
56
I15
86
58
-
18.5
14.5
10.5
19.5
15.5
11.5
17.5I 19.c
1 3 . 1> 15.c
1 OS1 11.c
20.5
16.0
12.0
1.1
is bossd
face ar*o.
cared velocities ond is mea,13.5
wed in incher of water.
11.0
8.5
G 7. Meorure
ceiling height in the
CLEAR only. This is the distonca
from the floor lo the towes, <oil12:o
ing beam or obstruction.
9.0
17.0
13.5
10.5
3.6
lo-
4 . Divergent Blow
blow, and cfm. The octuol meoswed c e i l i n g h e i g h t m u s t b e
equal t o o r greater t h a n t h e
minimum ceiling height for the
selection mode. Preferably the
top of on outlet should be not
ICII t h a n t w i c e t h e
outlet’.
height below the minimum ceiling height.
11.0
11.0
9. CfmlS.
Ft Outlet Wall Area
is thi stbndord for judging total
room air movcmcnt. T h e moxi-
a&&d thot furniturc,‘peopla,
etc..
obstruct 10% of the room
w&-section.
If I’& obstructions
vary w i d e l y f r o m IO%, the
“.I”.I
o
f
the cfm/sq ft oY1l.t
wall orco should be tampered
accordingly.
21.0
16.5
12.0
IO.
I
upward
U n d e r b l o w . I t is n o t olwoy,
nscasrary t o b l o w t h e anpi,.
length of the room unl.,,
there
a<* hoot l o a d sovrce, ot t h a t
a n d , squipmant
load, open
doors, sun-glass, etc. Considering
the concentration o f r o o m hoot
load on ths basis of Btu/(hr) (rq
ft), the outlet blow rhould cover
75% of the heat load.
-
15.5
12.5
9.5
FACTOR
I
3.
15.5
12.5
9.5
119
18.C 1 19.:
09 14.c 1 15.c
60 10.: i 11.0
I-
oh
indicoter dktonca from out.
let lo Iha p o i n t where Ihs air
stream
is substantially
dirsipatsd.
13.0
10.5
8.0
9 4 14.5
7 0 11.5
47 9.0
102 15.5
7 7 12.5
51 9.5
17.0
13.5
IO.0
the
2. Blow
8.0
13.0
10.5
8.5
_
deflect
word the celling.
I-
11.5
9.5
7.5
10.5
9.5
7.5
15
to
/np
Tet
15
20
2s
(fl)
M lir I CI( H t
- - - 58 10.5 11.0 12.0
44
8.5
9.0
9.5
29
7.0
7.0
7.5
a SIOW
10.0
8.5
7.5
Max
I. W h e n e m p l o y i n g ratings ,o,
flat ceiltngs, i t i* undsrrtaod
that the front IOUYI.I ore ,et
Str B = 1 . 3 6
Air
Puon.
tity
(cfm)
NOTES:
F o r appticotions
reqetring
o
limiting sound level-the out.
1st v e l o c i t y is l i m i t e d b y t h e
sound generotcd by the outlet.
2-W
!'.\Kl'
L'.
\I11
I~ISl‘l~ll~~J'I‘lON
/
TABLE
2I-WALL OUTLET RATINGS, FOR COOLING ONLY (Cont.)
For Flat Ceilings
OUTLET
VELOCITY
STATIC PRESSURE
STANDARD OUTLET
S T A T I C P R E !SSl I R E W I T H
METERING PLATE
Nom. Size
of Outlet
(and Free
Area)
VlYn.3
Setting
250 FPM
375 FPM
I
str B = .Ol, 2 2 % ” = .Ol
4 5 ” = .Ol
I
str E = ,013, 2 2 % ” = , 0 1 5
4 5 ” = .019
500 FPM
S t r B = . 0 1 , 2 2 % '= . 0 1 5 1 S t r B = . 0 2 4 , 2 2 '?O
/ = , 0 4 3 I str E = .061. 22x0 = .082
450x tl?R
I
45’ = .065
45” = .I 1,
I
Air
C luanmy
I
lcfm)
T e m p Dil
Fpi?
E
M
Air
(F) Quan.
_
25 t i t y
_
i
CIS H t
9.0 9 . 5
- fHow ‘”
(f0
(cfm)
12 x 8
(56.7)
113
16x a
(77.1)
Straight
22x0
45"
155
8.0
6.0
4.0
9.0
8.0
7.5
-
9.5
8.5
7.5
10.0
9.0 231
8.0
20 x 8
(97.6)
Straight
22'hO
45O
192
8.5
6.5
4.3
9.5
8.5
7.5
-
10.0
9.0
8.0
-
10.5
9.5 287
8.0
24 x 8
(118.0)
Straight
22'/2O
45O
231
9.0
6.9
4.5
9.5 10.0
8.5
9.0
7.5
8.0
-
30 x 8
(149.0)
Straight
22'/2O
45O
209
9 . 5 1I O . 0
7.0 9.0
4.7
8.0
10.5
9.5
8.0
36 x 8
(179.0)
Straight
221/2O
45O
350
9 . 9 1I O . 0
7.5 9.0
5.0
8.0
11.0 11.5
9.5 10.0
8.5
9.0
16 x 10
(97.7)
Straight
22x0
450
198
9.a
7.1
5.0
95
9.0
8.0
-
Ki
9.5
8.5
-
2 0 x 10
(124.0)
Straight
22'/i0
45O
249
10.5
8.0
5.2
Il.0
10.0
8.5
-
2 4 x 10
(150.0)
Straight
22%"
45O
300
11.0
8.4
5.5
1'0.5
9.5
8.0
1I I . 0
1 0.0
9.0
-
12.0
10.5
9.0
-
21.012.5
16.0 11.0
10.5 9.0
3 0 x 10
(195.0)
Straight
22x0
45O
364
12.0
9.0
6.0
I' 2 . 0
1 0.5
9.5
-
:2.5
Il.0
9.0
22.013.5
16.0 11.5
11.0 9.5
3 6 x 10
(227.0)
Straight
22'/20
45O
453
12.4
9.1
6.1
I 2.0
1 0.5
9.5
13.0
II.5
9.0
16 x 12
(118.0)
Straight
22x0
450
244
Il.0 I
8.1 1
5.5
20x12
(150.0)
Straight
22%"
45O
307
24 x 12
(181.0)
Straight
22x0
45"
30x12
(228.0)
36 x 12
(275.0)
a.s
7.5
7.0
-
8.0
7.5
,
lowTen
E
(fl)
_1J_
M
, Diff (F)
zig
22.0
16.0
Il.5
463
40 13.5
30 11.0
20 9.0
14.515.5
12.012.5
9.510.0
16.0 10.5
12.0 9.0
8.0 8.0
24.0
18.0
12.5
575
43 14.0
32 11.5
22 9.5
15.516.5
12.513.5
10.0 10.5
10.5
9.5 346
8.5
17.0 10.5
13.0 9.5
8.5 8.0
25.0
19.0
13.0
-
692
45 14.5
34 12.0
23 9.5
16.0 17.0 ,
13.0 14.0
10.010.5
11.0
10.0 435
8.5
18.0 11.0
13.0 9.5
9.0 8.5
26.0
19.0
13.5
868
46 15.5
35 12.5
23 10.0
17.018.0
13.515.0
10.511.0
18.0 11.5
13.0 9.5
9.0 8.5
27.0
20.0
14.0
1048
48 16.0
36 13.0
24 10.0
18.0 19.0
14.015.0
10.5 11.5
18.0 11.5
13.0 10.0
9.0 8.5
27.0 13.0 14.0 15.5
20.0 11.0 12.0 12.5
14.5 9.0 9.5 10.0
-
595
48 15.5
36 13.0
24 10.0
18.019.0
14.515.0
10.5 11.5
29.0
22.0
15.0
13.5 15.0 16.0
12.0 12.5 13.5
9.5 10.0 10.5
746
51 17.0
38 13.5
26 10.5
18.520.0
15.016.5
11.012.0
30.0
22.0
15.5
14.5 16.0 17.0
12.0 13.0 14.0
10.0 10.5 11.0
a99
55 18.5
41 14.5
28 11.0
20.021.0
15.517.0
12.012.5
14.516.0
12.0 13.0 751
10.010.5
32.0
24.0
16.5
-
15.0 17.0 18.5
13.0 14.0 15.0
10.0 10.5 11.5
1126
58 19.5
44 15.5
29 11.5
21.523.0
17.018.5
12.513.5
22.014.0
16.0 12.0
11.0 9.5
15.016.0
12.5 13.5 904
10.010.5
33.0
25.0
17.0
15.0 17.5 19.0
13.0 14.0 15.5
11.0 11.0 12.0
1355
60 20.0
45 16.0
30 12.0
22.023.5
17.519.0
12.513.5
21.012.5
16.0 11.0
11.0 9.0
13.515.0
11.5 12.5 488
9.510.0
Il.0
23.0
16.0
14.5 16.0 17.0
12.0 13.0 14.0
10.0 10.5 11.0
733
55 18.5
41 14.5
28 11.0
20.021.0
16.017.0
12.012.5
918
60
45
30
2o.c
16.c
12.C
22.0 23.5
17.5 19.0
12.513.5
1110
64 21.1
48 17S
32 12.C
24.025.0
18.520.0
13.514.5
1388
68
51
34
23.c
18.C
13.c
26.027.5
20.021.5
14.015.0
1673
7124.:
53 19.c
36 13.C
27.529.0
21.0 22.5
14.515.5
525
13s
1 1.c
9.c
-
370
13.0
10.0
6.5
13.:
12.c
9.:
-
Straight
221%"
45O
462
13.9
10.0
7.0
14.!
12.:
10s
Straight
22%"
45O
560
14.5
11.0
8.0-L
-
15.:
13.c
1o.c
I
8.7
GG
14.0 9.5
10.0 8.5
7.0 7.5
I
13.514.0
11.012.0
9.0 9.5
170
12.1
9.1
6.0
29.0
Air
Temp Diff (F)
( auanTi-pipr
tity
Min Clg Ht
(cfm)
339
8.5
7.5
12.5
11.0 367
9.5
M i n C f m / S qF t
O u t l e tW a l l A r e a
et)
str 8 = .175, 22%” = .I9
45’ = .27
36 12.0
27 10.0
18 8.5
12.0
IO.5
9.0
M a x C f m / S qF t
O u t l e tW a l l A r e a
e IlOH
&
M
Straight
22%"
45O
750 FPM
Str B = ,024, 22%" = .028
4 5 ” = ,035
15.0
1II.0
13.0 14.0 15.5
11.0 12.0 12.5
9.0 9.5 10.0
14.0
12.0 460
9.5
22.0 14.c 15.016.0
16.0 12.C 12.513.5
11.0 9.2 10.010.5
613
33s
25.C
17.c
16.0 17.5 19.0
13.0 14.0 15.5
1 1 . 0 1 1 . 01 2 . 0
14.7
12.5 555
10.0
24.0 14.: 16.018.0
18.0 12.C 13.0 14.0 740
1 2 . 0 1 o . c 10.0 11.0
35.c
26.C
18.C
-
17.0 18.5 20.0
14.0 15.0 16.0
11.0 11.0 12.5
15.5
13.0 695
10.5
25.0 15.: 17.018.0
19.0 13.c 14.015.0
12.0 1o.c 10.511.0
37.c
28.C
19.c
18.0 20.0 21.5
14.5 16.0 17.0
16.5
14.0
10.5
27.0 16.C
20.0 13.5
13.0 10.5
836
t
K
925
39.c
29.C
2o.c
-
FACTOR
19.0
14.0
9.6
5.7
4.2
2.9
(;~-I.\I’IxR
3. itoobr .\III I)Is’r‘1~II:II’r.IoN
2-81
TABLE 21-WALL OUTLET RATINGS, FOR COOLING ONLY (Cont.)
For Flat Ceilings
1000 FPM
OUTLET
VELOCITY
~._..
- - STATIC PRESSURE
T A N-D A R D O U T L E T
- S___--~
STATIC PRESSURE WITH Str B = .33, 22'/2" =
METERING PLATE L
- r
Nom. Size
Air
of Outlet
Vane
C2uan- e
(and Free
Setting
tity
Area)
(cfm)
1500 FPM
-.__....~
S t r B = , 2 1 1 , 22%" =
-
2000 FPM
.24
StrB = , 3 7 5 , 2 2 % " = . 4 2
45" = ,565
B = ,715, 22X0= .74
45O = 1.15
.33
# D i f f (F)
. E
201
Clg Ht
12 x 8
(56.7)
Straight
22%"
45"
452
16 x 8
(77.1)
Straight
22%"
45"
616
20 x 8
(97.6)
Straight
22%"
45O
770
62 16.0
47 13.0
31 10.0
?4 x a
18.0)
Straight
22x0
450
920
65 17.0
49 13.5
33 10.5
19.020.0
15.0 16.0 1384
11.012.0
30 x 8
(149.0)
Straight
22'/z0
45"
1160
60 17.5
51 14.0
34 10.5
19.521.0
15.5 17.0 1736
11.512.5
36 x 8
(179.0)
Straight
22x0
45O
1404
71 18.5
53 14.5
36 11.0
16x10
197.7)
Straight
22%O
45O
20x10
(124.0)
15.016.5
12.513.5
10.010.0
18.0 19.0
14.015.5
10.511.5
678
16.518.0
1 3 . 5I
1 4 . 5i926
10.01
1
l.O!
I . W h e n e m p l o y i n rg a t i n g s ‘ o r
Str B = 1.36
flat
Blov
(ft)
904
121
91
62
TY
I
Temp Diff (F)
I5~
2.
p5
18.0 20.0 21.5
14.0 15.5 17.0
10.5 11.5 12.0
13a
100
67
21.023.0
15.5 18.0
12.013.0
1540
144
108
72
21.0 23.5 26.0
16.5 18.0 20.0
12.0 13.0 14.0
1 0 72 0 . 0
80 15.5
54 11.5
22.524.5
16.518.5
12.513.0
1840
151
113
76
22.0 25.0 27.5
17.5 19.0 20.5
12.0 13.5 14.5
111 21.0
83 16.5
56 11.5
t
23.525.5
17.519.5
13.013.5
2320
157
118
79
23.5 26.0 29.5
18.0 20.0 21.5
13.0 14.0 15.0
20.522.0
16.0 17.5 2096
12.012.5
116 21.5
87 17.0
58 12.0
25.026.5
18.020.0
13.014.0
2808
164
123
82
24.5 27.5 31.0
19.0 21.0 22.5
13.0 14.5 15.5
792
71 18.0
53 14.5
3611.0
20.522.0
16.0 17.5 1190
12.012.5
116 21.0
87 17.0
58 12.0
25.026.5
19.020.0
13.014.0
1584
164
123
82
Straight
22%"
450
994
75 19.5
56 15.5
38 11.5
22.024.0
17.0 18.5 1492
12.513.5
1 2 22 3 . 0
92 18.0
61 13.0
26.529.0
19.021.5
14.015.0
1988
174
130
a7
2 4 x 10
(150.0)
Straight
22Yi"
45"
1200
a021.0
60 16.5
40 12.0
24.025.5
18.5 19.5 1798
13.014.0
<l 24.5
98 19.0
66 13.5
28.531.0
20.022.5
14.515.5
2400
185
139
93
3 0 x 10
(195.0)
Straight
22%"
45"
1502
25.527.5
19.5 21.0 2252
14.015.0
139 26.5
104 20.5
70 14.0
30.534.0
21.524.5
15.516.5
3004
196
147
98
36x 10
(227.0)
Straight
221/2O
45O
I808
87 23.5
65 18.5
44 13.0
26.528.5
20.0 21.5 2710
14.515.0
142 27.0
106 21.0
71 14.5
31.535.0
23.025.0
16.017.0
16 x 12
(118.0)
Straight
22%O
45O
976
81 21.0
61 16.5
41 12.0
24.025.5
18.5 19.5 1466
13.014.0
131 24.5
98 19.0
66 13.5
28.031.0
21.522.5
14.515.5
-0x12
(150.0)
Straight
22w
45O
1226
142 27.0
106 21.0
71 14.5
2 4 x 12
(181.0)
Straight
22x9
45O
1480
153 29.0
115 22.0
77 15.0
34.0 37.5
25.026.5
17.0 18.0
30x12
(228.0)
Straight
22'/2O
45O
1850
163 31.5
122 24.0
8216.0
36.5 40.5
27.0 28.5
18.0 19.0
36x 12
(275.0)
Straight
22Yi"
45"
2230
172 33.5
129 25.0
0616.5
38.042.5
28.030.0
18.5 20.0
1 6 . 0 /1 7 . 5 1
-L
K
+
Blow mdicoter
Min Clg Ht
1232
t
ceilings, it is undwrteod
the front louvrer
are re,
to deflect the air upword
toward the ceiling.
that
Air
Quantity
(cfm)
19.521.0
14.017.0
11.512.0
43 12.5
29 9.5
NOTES:
dirtonts
from out-
pat-d.
20.0 22.0 24.0
15.5 17.0 18.5
11.0 12.0 13.0
4.
Divergent
S l o w ha vertical
lowres
Itmight
forward in the
center, with uniformly increasing
angular d e f l e c t i o n 10 a m a x i mum 01 each end. The 45” div e r g e n c e signifisr a n a n g u l a r
dsflection 01 each e n d o f t h e
outlet o f 450, and similarly for
22%0 divsrg.ma.
Velocity is
foes area.
bored
on
effective
Measure ceiling height in the
CLEAR only. This is the dirtonce
from Ihs floor to the lowest ceiling beam or obstruction.
The Minimum Ceiling Height
(table) is the minimum ceiling
height which will
give proper
operation of the outlet for the
given outlet velocity, vane setting,
temperature difference,
blow, and cfm. The odwl mcorwed c e i l i n g h e i g h t m u s t be
equal to 01 g r e a t e r t h a n t h e
minimum ceiling height for the
selection
made.
Preferably
the
top of an outlet should
be not
less than t w i c e t h e outlet’s
height below the minimum cciling height.
28.0 32.0 36.5
2 1 . 5 I2 4 . 0I 2 5 . 5
1 4 . 5 /1 6 . 5 !1 7 . 0
31.535.0
I’21 36 .. 50 21 57 .. 00
9.
Cfm/Sq
Ft Outlet Wail
Area
is the rtondard
for judging total
room air m o v e m e n t . The m a x i m u m value shown results i n an
air mwement in Ihe zone of ocCUPD~CY
o f about 5 0 fem. It i s
aswmeb
t h a t furniture,‘peopls,
etc., obstruct IO% of the room
UOII-section. If the obstructions
vary w i d e l y f r o m
IO’%. t h e
values o f the cfm/rq
f t outlet
w a l l area should
be tempered
accordingly.
10. For applicqtionr requiring a
limiting round level--the out.
FACTOR
Max Cfm/Sq F t
O u t l e tW a l l Area
7.2
4.8
3.6
Min Cfm/Sq Ft
Outkt Wall Area
2.2
1.4
I.1
l e t v e l o c i t y i s l i m i t e d b y the
sound generated by the outlet.
TABLE 21-WALL OUTLET RATINGS, FOR COOLING ONLY (Cont.)
For Beamed Ceilings
I
OUTLET
STATIC PRESSURE
STANDARD OUTLET
STATIC PRESSURE WITH
METERING PLATE
Nom. Size
of Outlet
(and Free
Area)
375
250 FPM
VELOCITY
Str B = .024, 22%" = .043
450= .065
str B = .OI, 22'/b0 = 015
450= 03A .
-
8x4
(16.91
Straight
22Yl"
45O
10 x 4
(21.7)
Straight
22%"
45O
3.5
2.5
1.8
8.2
7.6 57
7.0
f
7.4
5.5
3.7
8.2 8.8
7.5 8.0
6.9 7.2
9.2
8.2
7.5
12 x 4
(24.6)
Straight
22'/rt0
4s"
3.5
2.5
1.8
8.2
7.6
7.0
7.5
5.5
3.9
-
8.3 8.8
7.6 8.0
7.0 7.2
16 x 4
(35.9)
Straight
22%"
45O
20 x 4
(45.5)
Straight
22%"
45O
7.9
6.0
4.0
8.0
6.0
4.0
-
24 x 4
(55.0)
Straight
22J/l"
45O
93
30 x 4
(68.3)
Straight
22x0
45O
36 x 4
(83.5)
Straight
22Yl"
45O
8x6
(26.5)
Straight
221x0
45O
5:
10 x 6
(34.0)
Straight
22x0
45O
6t
12 x 6
(41.6)
Straight
22%"
45O
a(
16 x 6
(56.6)
Straight
22x0
45O
10;
20 x 6
(71.5)
Straight
22x0
45O
5
3
2
24 x 6
186.5)
Straight
22%"
45O
3
1
5
30 x 6
(109.0)
36 x 6
(131.3)
61
emp Diff (IF)
t5
%-Kg t
str B = .
061I 22'/z0 = .082
45O= ,118
Air
cavan.
tity
(cfm)
8.1
7.5
7.0
8.2 8.7
7.5 7.9
,5 . 5 6 . 5
44
68
3.7
2.7
2.0
10.0
7.5
5.0
(cfm)
I-
9.3
0.4
7.5
tity
9.8
a.8
a9
7.0
17.0
13.0
9.0
9.7
8.5
7.4
9.2
8.3
7.5
91
11.0
8.1
5.5
8.9
8.1
7.2
9.4 9.9
8.5 8.9
7.67.9
136
18.0 9.9
13.0 8.6
9.0 7.5
8.4 8.9
7.7 8.1
7.0 7.3
9.4
8.4
7.6
122
11.0
8.1
5.5
9.0
8.2
7.3
9.610.0
8.6 9.0
7.77.9
la3
19.010.0
14.0 8.8
10.0 7.6
8.4 9.0
7.8 8.2
7.07.4
9.4
8.5 154
7.6
11.5
8.5
6.0
9.1
8.3
7.4
-
9.6 10.1
8.7 9.1
7.8 8.0
231
20.010.2
15.0 8.9
10.0 7.6
11.5
8.5
6.0
9 . i,
8 . 3I
7 . 4I
9.610.2
8.8 9.1
7.8 8.0
278
20.010.2
15.0 8.9
10.0 7.6
9.:
8 . LI
7 . 4I
9.710.3
8.5 9.0
7.88.2
7.17.4
9.5
8.6
7.6
116
4.2
3.1
2.1
8.5
7.9
7.2
175
8.0
6.0
4.0
-
8.5 9.0
7.9 8.3
7.17.5
9.6
0.6 233
7.7
12.0
9.0
4.6
i4a
4.4
3.3
2.2
-
8.5
7.9 210
7.2
8.0
6.0
4.0
8.5 9.1
7.9 8.3
7.17.5
9.6
8.6 279
7.7
12.0
9.0
6.0
-
5 . c)
3 . c3
2 . :i
9.0
8.3
7.4
77
9.5
7.0
4.8
9.0 9.6
8.3 8.8
7.37.8
1IO.2
:
9.2
103
8.0
5.15
4 . 1I
2 . fI
__
6.c1
4 . :5
3.c)
9.6
8.7
7.8
98
9 . 6 1 0 . 2'1 1 0 . 9
8.7 9.3
9.8
7.6 8.0
8.4
ia5
131
1 l . CI 9 . 7 1 0 . 4 I 11.1
8.8 9.4
9.8 159
8.1
8.5
5 . 51 7 . 7 8 . 1
1 2 s) 1 0 . 1 1 0 . EI 1 1 . 6
9s)
9.1 9.7 10.2 214
6 . CI
7.9 8.4 I
8.8
(F)
15 i 20 1 25
Mill Clg Ht
18.0 9.8
13.0 8.6
9.0 7.4
8.0
6.0
4.0
Diff
tft)
112
139
114
in Clg HI
8.7
8.0
7.2
i
8.4
7.8
7.1
,
.t:, _ Blow Temp
75
4.1
3.1
2.0
,
%p Diff (F)
1 20 / 25
9.4 9.9
8.5 8.8
7.6 7.0
115
I::
7.8
str B = 1
. 7 5, 2 2 ' / 1 O = . I 9
45" = .27
8.8
10.5
8.0 8.0
5 ._
4
7.2
_
8.4
7.8
7.1
-
FPM
str B = ,051 22'/Iz0 = ,061
45O'= .oa
I
Air
hmn
tity
(cfml 3.5
30
2.5
1.8
Vane
setting
750
500 FPM
FPM
S,r rJ = .Ol, 22x0 = .Ol -- str B = ,013, 22'Lz0 = .015 Str 8 = ,024 22'/1' = 028
45-o; .035 .
45O= .Ol
45" = ,019
1
21.010.3
--I-
9 . :3
9 . 1I
7 . 1I
I L 21 2 . 0
9.7 10.3
8.28.6
12.2 13.1
10.411.2
8.6 9.2
13.0 7?
1 0 . 0 8 . 13
6.0
7 . ;7
15.0
11.0
7.0
1 0 . 5:
9 . :3
8 . 'I
13.314.2
11.2 12.1
9.1 9.7
15.0
11.0
7.0
1 0 . 61
9 . 24
8 . 'I
13.614.5
11.412.2
9.2 9.8
1 2 . c1
9 . c1
6 . ti
1 0 . 5 1 1 . :, 1 2 . 1
9 . 3 1 o . cI 1 0 . 5 2 6 9
8 . 1 8 . C,
9.0
1 6 . 0 i - L2
12.0
9.8
8.0 a . ,4_ :
17.a
13.0
9.0
8 . 6 9 . 2 /9 . 6 1
10.7
9.5 24:
8.5
1 3 s1
1 o . c1
6 . :5
1 0 . 7 1 1 . :i 1 2 . 3
9 . 5 1 0 . :t 1 0 . 7 3 2 4
8 . 2 a . 13
9.2
18.0
13.0
9.a
15.716.9
12.913.9
10.210.8
Straight
22%"
45O
11.0
9.7 30'
8.6
1 3 . cI
1 o . c1
6.5
1 1 . 0 1 1 . tI
9 . 6 1 0 . ~1
8.4 0.9
I 9.0
14.0
10.0
16.117.4
13.314.2
10.4 11.1
Straight
22x0
45"
11.1
9.8
8.7
1 3 . 0 '1I l . 21 2 . 0 1 2 . 4
10.0
9.810.6 '1 Il.1
6.5
8.5 9.1
9.5
-
19.0
14.0
10.0
-
16.618.0
13.514.6
10.611.3
2
7
2
10.2
9.0
8.1
161
368
K
12.6
10.9 406
9.4
490
14.515.5
12.012.9
9.710.3
1 1 6 . 0 19 . 3
FACTOR
Max Cfm/Sq
outlet
Wall A Ft
rea
29.0
19.0
14.0
9.6
OM ui tn lCeftWma/lSlqFAtr e a
8.7
5.7
4.2
2.9
15.216.3
12.513.5
10.010.6
TABLE 21-WALL OUTLET RATINGS, FOR COOLING ONLY (Cont.)
For Beamed Ceilings
OUTLET
VELOCITY
2000 FPM
jtr 6 = . 3 7 5 , 2 2 % " = . 4 2
A S ” = SAS
STATIC
PRESSURE
STANDARD OUTLET
S T A T I C PRti s s U R E WIT1
METERIt 4 G PLATE
L
Nom. Size
Air
o f Outlei
V a n e
cauanS
e
t
t
i
n
g
tity
(and Free
Area)
(cfm)
8x4
(16.9)
Straight
22%"
45O
118
10 x 4
(21.7)
Straight
22s"
45O
150
12 x 4
(24.6)
Straight
22%"
45O
181
Straight
22%"
45O
244
Straight
22s"
45O
24 x 4
(55.0)
Straight
22%"
45"
30 x 4
(68.3)
Straight
22'h0
45O
Str 8 = 1.36
Air
D i f (F)
B l0WTBm
- Qucln.
2
5
(it) 15 2 0
_ tity
Mi
C l ? It 4t
(cfm)
24 10.4
1.2 1 1 . 9
18 9.0
9.6 10.2
177
12
7.6 8.0
8.4
f:
26 10.6
19 9.1
13 7.7
Il.4
9.7
8.2
12.1
10.4
8.5
27 10.7. 11.6
20 9.3
9.9
14 7.8 8.2
2 8 11.0 1 1 . 9
21
9.5 1 0 . 1
14 7.8
8.4
29 11.2
22 9.7
15 7.9
466
237
299
60 13.0
45 10.4
30 8.0
14.2 15.4
11.3 12.0
8.5 8.9
362
62 13.3
47 10.6
31 8.2
14.5 15.7
11.6 12.3
8.7 9.2
65 13.7
4 9 11.0
33 8.5
15.0 16.2
12.012.8
9.0 9.5
67 14.0
50 11.3
34 8.6
15.316.6
12.3 13.1
9.29.7
68 14.2
51 11.4
34 a.7
15.516.8
12.413.3
9.3 9.9
70 14.4
53 11.6
35 8.9
15.717.0
12.613.5
9.5 10.0
272
44 12.3
33 10.2
22 8.0
13.4 14.4
1 0 . 9 11.6
8.5 8.9
12.7
10.7
8.6
366
46 12.6
35 10.4
23 8.2
13.8 14.8
Il.3 12.0
8.7 9.2
12.1
10.3
8.5
-
12.9
10.9
8.8
462
48 12.9
36 10.6
24 8.3
12.3
10.4
8.5
12.4
10.5
8.6
-
13.1
11.0
8.8
556
49 13.0
37 10.8
25 8.4
14.2 15.3
I 1 . 6 12.4
9.0 9.4
13.3
11.2
8.9
698
50 13.2
37 10.9
25 8.5
14.4 15.5
11.8 12.6
9.1
9.6
932
51 13.3
3 8 11.0
26 8.5
14.5 15.7
11.9 12.6
9.2 9.6
1116
206
36 11.6
2 7 10.8
18
8.6
Straight
22s"
45O
262
13.8
11.5
9.2
1 5 . 2,
12.6 ,
9 . 5,
12 x 6
(41.6)
Straight
22'h0
45O
16 x 6
(56.6)
Straight
22Yi"
45O
‘-Ax6
(71.5)
Straight
22%"
45O
24 x 6
(86.5)
Straight
22%"
45O
14.9
12.3
9.6
310
I-
M i n Clg Ht
5 8 1 2 . 7 (1
44 10.2
29 7.8
12.3
10.5
8.6
Straight
22%"
45O
16.217.5
13.114.0
10.0 10.6
15.917.2
12.813.6
9.6 10.1
412
82 16.2
62 13.0
41
9.9
17.919.3
14.215.3
10.611.3
16.5
13.5 392
10.4
66 16.1
50 13.2
3 3 10.1
17.9 19.5
14.4 15.5
10.8 11.6
S24
9 2 18.1
69 14.3
46 10.8
20.0 21.7
15.717.0
11.612.5
318
15.5 , 1 6 . 9
12.8 1 3 . 7 4 7 6
10.0 / 1 0 . 5
67 16.4
50 13.4
34 10.2
18.319.9
14.715.7
11.0 11.8
636
94 18.5
70 14.6
4 7 11.0
20.4 22.1
16.017.3
11.812.7
428
16.4 1 8 . 1
1 3 . 6, 1 4 . 6
10.5 1 1 . 1
-
642
72 17.6
54 14.3
36 10.8
19.7 21.3
15.6 16.9
11.6 12.5
856
102 20.0
77 15.6
51 11.6
22.123.9
17.118.7
12.513.5
19.0
15.3
11.5
806
77 18.6
58 14.9
39 11.2
20.822.4
16.417.8
12.113.0
1076
108 21.1
81 16.4
54 12.1
23.325.2
17.919.7
13.0 14.1
19.7
15.7
11.8
972
79 19.3
59 15.4
4 0 11.6
21.623.2
16.918.4
12.413.3
1296
111 21.9
83 17.0
56 12.4
24.2 26.1
18.520.4
13.414.5
82 20.0
62 15.9
41 11.7
22.424.0
17.419.0
12.613.7
1624
115 22.7
86 17.5
58 12.7
25.027.1
19.1 21.1
13.714.9
84 20.6
63 16.3
4 2 11.9
23.024.8
17.819.6
12.914.0
1960
119 23.4
89 18.0
60 13.0
25.828.0
19.621.7
11145..03
17.5 ,
14.2
10.8
1 8 . CI
14.6 ,
11.1
648
Straight
22x0
Straight
22%Q
45"
(cfml
2.913.8
0 . 5 11.2
8.1 a.5
3.1 14.1
0.7 11.4
a.48.7
12.5
10.6
8.6
-
36 x 6
(131.3)
:Ig H t
42 12.1
3 2 10.0
21
7.8
31 11.5
2 3 10.0
16
8.1
30 x 6
(109.0)
Min
224
Straight
22'h0
45O
8x6
(26.5)
Air
auan.
tity
18.7 20.4
15.c I 1 6 . 2 1 2 1 8
11.4 I 1 2 . 1
19.; I
15.4
1 I.6 ,
K FACTOR
980
Max Cfm/Sq F t
o u t l e t W a l l Area
7.2
4.8
3.6
M i n Cfm/Sq F t
Outlet Wall Area
2.2
1.4
1.1
i
t
NOTES:
I.
angular daflsction io a m a x i mum at eorh end. The 45” divetgsnca
rignifias an angular
deflection ot sach end of Ihs
outlet of 45’. and ~imilorly for
22’h” divergenca.
3 . Underblow. I t i, n o t olwoy,
nace,sory t o blow tha entire
length of the roam unless there
or., heat load sourc.,~ a t
that
end, squipmsnt load, open
door,, wn-glars, *tc. Gmidering
the toncsntralion o f room heat
load on the basis of Btu/(hr) (q
ft), the outl*t blow should cov*r
7536 of the hsat load.
4. Velocity is based on effective
face area.
6. Measure ceiling height in the
CLEAR only. Thb is the distance
from the floor to the low*s1 c*iling beam or obstruction.
7. The Minimum Ceiling Height
(table) is ths minimum csiling
haighl which will givs proper
operation of the wtl*t for the
given outlet v&city. van* s*tii.%&
tempsroture
difference,
blow, and cfm. The actual mea+
wed ceiling height must be
equal t o or g,.mtc, t h a n t h e
minimum ceiling height for the
rsledion mad*. Preferably the
top of an outlet should be not
less t h a n twice
the o u t l e t ’ s
height below the minimum ceiling height.
8. Cfm/Sq Ft O u t l e t W a l l A r e a
is the standard for judging total
room air movement. The maximum value shown results in an
air movement
in the zone of oc.
cupancy
of about 50 fpm. It is
orrumed that furniture. people,
etc., obstruct IO% of the room
cross-section. If the obrtructionr
vary w i d e l y f r o m IO%, the
values
of the cfm/sq h outlet
wall orso should
be tempered
accordingly.
9. For opplicationr requiring a
limittng round level-the outlet velocity is limited by the
sound gencr.tcd by the outlet.
I’\K I ‘
2-84
2. .\IK I~IS’1‘1~11117’1~10N
TABLE 21-WALL OUTLET RATINGS, FOR COOLING ONLY (Cont.)
For Beamed Ceilings
OUTLET VELOCITY
STATIC PRESSURE
STANDARD OUTLET
S T A T I C P R E S S U RWEI T b
METERING PLATE
Nom. S i z e
ofOutlei
(andFree
Area)
250 FPM __-__.
str B = .O', 22x0 = .Ol
45" = .O'
375 FPM
500 FPM
750 FPM
I
45" = .0'9
S t r B = . O l . 2 2 % " = , 0 1 5 1S t r B = , 0 2 4 , 2 2 % ' = . 0 4 3
45"'= ,028
Diff ( F )
-C
Valle
Setting
5irpc
12 x 8
(56.7)
Straight
22%"
45"
2.''3.0
0.7'1.2
9.2 9.8
16 x 8
(77.1)
Straight
22%"
45"
3.114.0
1.411.9
9.7 10.1
Clg Ht
I
81
(
(
Min Clg Ht
T-;
1
1
36
27
'8
'5.4
12.6
10.0
17.0 18.3
13.7 '4.8
10.8 11.4
2
1
1
1 5.0 '6.1
1 2.7 13.7 463
1 0.4'0.9
40
30
20
16.9 '8.6 20.3
13.7 15.0 '6.2
'0.6 11.5 '2.2
Straight
22x0
45"
1 9 2
3.8'4.9
1 . 91 2 . 6
0.0 10.6
385
2
1
1
1 5.9 '7.2
1 3.414.5
I 0.811.5
57s
43
32
22
18.'
14.5
11.3
20.0 21.7
'6.0 17.3
12.1 13.0
24 x 8
(118.0)
Straight
22x0
45"
231
14.3 15.6
12.3 '3.2
10.2 10.8
460
2 ! 5 . 01 5 . 2
1 9.012.8
1 3.010.4
1 6.6 18.1
1 3.9 15.1
1 1.0 '1.8
692
45
34
23
19.0
15.2
11.6
21.0 22.7
16.8 '8.2
12.5 13.4
30 x 8
(149.0)
Straight
22x0
45"
209
14.9 '6.3
12.7 '3.7
10.5 11.1
580
:! 6 . 01 5 . 9
1i 9 . 01 3 . 4
1 3.510.7
1 7.319.0
I 4.515.7
1 1.4 12.2
868
46
35
23
20.0 22.0 23.9
16.0 '7.5 19.'
'2.0 13.0 '4.0
36 x 8
(179.0)
Straight
22s"
45O
18.0 '4.3 15.5 16.8
13.0 12.' 13.1 14.1
9.0 10.1 10.7 '1.4
702
:! 7 . 0' 6 . 5 1 i 8 . 01 9 . 7
:2 0 . 0 1 3 . 7 1 , 5 . 0 ' 6 . 2 1048
11 4 . 0 1 0 . 9 1 1I l2 . 67 1
16 x 10
(97.7)
Straight
22x0
45O
' 8 . 0 1 4 . 1 I ' 5 . 5 1 6 . eI
13.0 12.1 '3.1 '4.'
9.0 10.1 '0.7 11.4
396
27.0 16.3
20.0 13.7
14.5 10.9
20 xa
(97.6)
'8.0 13.8
'3.0'I.8
9.0 9.9
9.6 '2.2 13.3 14.2
7 . ' i1 0 . 8 I1 1 . 7 ' 2 . 5
5.0
9.3
9.8 '0.4
297
1 0 . 5 /1 3 . 1 !1 4 . 1
8.0 11.4 '2.3
5.2
9.7 10.2
'5.2
'3.2
'0.8
374
'6.1
13.9
11.2
t
t
I
11 8 . 0
19.7
Il.7
12.6
11 5 . 0 1 6 . 2
595
16.5 18.C
'3.8 15.C
'1.1 11.5
497
19.3 21.2
15.9 17.3
12.2 13.3
746
450
'7.6 '9.1
14.6 15.5
11.5 12.:
600
20.6 22.5
'6.8 '8.3
12.6 '3.8
099
17.3
14.8
11.7
564
1 8 . 9 2 0 . :i
1 5 . 5 1 6 . 1P
1 2 . 0 1 2 . e1
751
22.1
'8.0
13.3
24.4
1 9 . 6 1126
14.6
.
17.7
15.1
11.8
680
1 9 . 5 2 1 . 1I
1 5 . 81 7 . : I
' 2 . ' 1 3 . c1
904
22.7
18.3
13.5
25.0
20.0 1355
14.9
16.1
13.9
11.2
367
1 7 . 6 1 9 .I‘
1 4 . 6 1 5 . 1?
1 1 . 5 1 2 . :2
488
31.0'8.7
23.015.4
'6.011.9
Straight
22x0
45"
'7.7
15.1
11.8
460
2 2 . 0 1 7 . ; 1 9 . 52 1 . '
16.0 14.1 '5.8 17.:
11.011.1 '2.113.c
613
33.020.5
25.0'6.9
17.013.7
24 x 12
(181.0)
Straight
221/2=
45O
18.8
16.0
'2.2
555
24.0 18.; 20.723.1
18.0IS.< 16.8 '8.:
12.0 11.1 '2.613.:
30x12
(228.0)
Straight
22r/,O
45O
462
20.2
'6.9
12.7
695
36 x 12
(275.0)
Straight
22x0
45"
560
21.3
'7.6
'3.1
036
2 0 x 10
(124.0)
Straight
22x0
45O
2 4 x 10
(150.0)
Straight
22%"
45O
3 0 x 10
(195.0)
Straight
22x0
45O
3 6 x 10
(227.0)
Straight
221/2O
450
16 x 12
(1'8.0)
Straight
22vi"
45O
MClX ctm,sq rt
O u t l e tW a l l A r e a
Min C f m / S qF t
o u t l e tW a i l A r e a
249
364
244
-l-L
'4.5 18.0 '9.8
11.0 15.0 '6.3
8.0 11.5 '2.2
19.015.;
'4.0 12.5
25.02O.f
'9.016.,
22.224.t
17.719.2
13.2 14.
48
20.8 22.9 24.8
36
'6.4 '8.2 19.7
24 112.4/13.3/14.4
55
41
28
24.1 26.5 28.4
I1 8 . 8 I2 0 . 7 !2 2 . 4
13.7 14.8 16.1
925
23.6 25.:
18.5 20.:
13I
. 6 14~
K F ACl'OR
29.0
19.0
14.0
9.6
8.7
5.7
4.2
2.9
.
T A B L E 21-WALL
OUTLET RATINGS, FOR COOLING ONLY (Cont.)
For Beamed Ceilings
OUTLET
1000 FPM
V&OCITY
1500 FPM
2000 FPM
STATIC PRESSURE
STANDARD
_- .-..STATIC PRES
METERIN
Nom.
of
Size
Outlet
(and
Free
OUTLET
IRE
WITH
PLATE
Van*
Setting
Area)
--~-~-_-~
str 6 =
1
.33, 2 2 % ” =
.33
Str B
=
,715. 22%’ =
.74
str
i
Air
i
cfm)
Straight
12 x 8
(56.7)
~.-
22vi"
45O
16 x 8
(77.1)
Straight
20 x a
(97.6)
Straight
1x8
.1&O)
22fi"
45O
Straight
16 x 10
(97.7)
Straight
20 x 10
(124.0)
Straight
24
Straight
22s"
45O
22%O
45O
22%"
45O
221/2O
450
22'/i0
450
30x10
(195.0)
Straight
36x10
(227.0)
Straight
16 x 12
!118.0)
Straight
(150.0)
24 x 12
(181.0)
I.
678
904
926
1232
2 9
2. Blow indicotsr dirtoncs from outl e t t o ths p o i n t w h e r e the air
s t r e a m is rubstantiolly
dirripotad.
1
jl1.6/12.7/13.51
3.
Underblow. It
i s not alwoyl
nscer,ory
lo b l o w the e n t i r e
,eng,h of ths room unless there
ore heat load s o u r c e s ot t h a t
end, e q u i p m e n t l o a d , o p e n
doors, wn-gloss, etc. Considering
,he concentration of room hoot
load on the barb of Btu/(hrl (sq
f,), the outlet blow should
cover
7S% of the heot load.
4.
Velocity is
foes area.
I
1540
1
3 1 1 1 2 . 4 1 1 3 . 41 1 4 . 5 1
22'/i0
45O
22%"
45O
22x0
45"
33 12.9 14.0 15.1
I160
68 23.3 26.0 28.4
51 18.4 20.3 22.1 1736
34 13.3 14.5 15.7
2320
157 31.4 35.5 40.4
1 1 82 4 . 0 2 7 . 0 2 9 . 2
79 16.3 18.2 19.6
1404
71 24.3 27.2 29.5
53 19.0 21.2 22.8 2096
36 13.8 15.0 16.2
2808
164 32.0 37.3 42.2
1 2 32 4 . 8 2 8 . 1 3 0 . 2
82 16.8 18.8 20.2
792
7 12 3 . 9 2 7 . 2 2 9 . 5
53 19.0 21.2 22.8 1190
36 13.8 15.0 16.2
1584
164 32.3 37.3 42.2
123 24.8 28.1 30.2
82 16.8
18.8 20.2
994
75 26.0 29.6 32.0
56 20.6 22.7 24.6 1492
38 14.6 16.0 17.3
1988
30 x 12
1228.0)
Straight
36 x 12
(275.0)
Straight
22%"
45O
22'/i0
45O
35.3
40.2
46.6
130
27.0
30.3
32.7
87 17.9 20.3 21.7
1200
1798
1
I 2400
I
1502
86 30.2 34.6 37.4
65 23.6 26.3 28.2
43 16.3 18.2 19.3
2252
i
t
I
3004
147 31.1 35.4 38.0
98 20.1 22.8 24.5
87 31.3 36.0 38.5
65 24.2 27.0 29.0
44 16.6 18.5 19.7
2710
)
1 3616
I
200 43.0 49.1 55.8
1808
5
1952
158 38.1 43.4 50.0
1 3 82 8 . 6 3 2 . 5 3 4 . 8
93 18.8 21.5 22.8
185 38.1 43.4 50.0
139 28.6 32.5 34.8
93 18.8 21.5 22.8
196
41.7
47.0
54.3
150
32.0
36.4
39.2
100
20.5
23.3
25.0
976
81 28.0 32.0 34.2
6 12 1 . 8 2 4 . 2 2 6 . 1
41 15.3 16.0 18.1
1466
1226
87 31.3 36.0 38.5
65 24.2 27.0 29.0
44 16.6 18.5 19.7
1836
1
1
I
2452
200 43.0 49.1 55.8
150 32.0 36.4 39.2
100 20.5 23.3 25.0
1480
93 33.6 38.6 42.0
70 25.8 28.9 31.0
47 17.4
19.5 20.8
2220
5
2
5
2960
213 47.0 53.2 59.2
1 6 03 4 . 2 3 9 . 2 4 2 . 1
107 121.7124.6126.5
1850
98 36.7 42.4 46.2
74 27.8 31.1 33.6
49 18.4 20.6 22.1
2776
Straight
22'h0
45O
174
80 28.0 32.0 34.2
60 21.8 24.2 26.1
40 15.3 16.9 18.1
Straight
22vi"
45O
1 a40
920
2230
I
I
3
3700
I
1
-7
4 4460
2
103 39.4 45.1 49.8
77 29.4 33.0 35.7
3346
5 2 1 1 9 . 1( 2 1 . 0 1 2 3 . 0 1
K
Divergent Blow hor
verlicol
,ouvr., rtroight forword
i n the
ccnkr,
with uniformly increasing
clngular d e f l e c t i o n lo o moximu,,,
a, each e n d . The 45’ divergence s i g n i f i e r o n ongulor
deflection of each end of Ihs
ode, o f 450, ond rimilorly for
22’b”
divergence.
1
Straight
36 x a
(179.0)
2 0 x 12
-
,375
1
22%"
45"
Straight
x 10
452
22%"
45O
30 x 8
(149.01
(150.0)
=
Air
luan- tl
tity
(
Icfm)
“on-
lily
B
hosed
on
effective
5. S,o,lc P,errure is t h a t pressure
r e q u i r e d lo p r o d u c e t h e indicored velocities o n d i s meos- I
wad in inches of w&w.
6.
Measure ceiling height in the
CLEAR only. This b the dirtonce
from the floor lo the lowest cciling beam or obstruction.
7.
The
Minimum
Ceiling
Height
(table)
is t h e m i n i m u m ceiling
height which will give proper
operation of the outlet for the
given outlet velocity, vans retting,
temperature difference,
blow, and cfm. The cc+u.l meoswed c e i l i n g h e i g h t
must b e
equd t o o r g r e a t e r t h a n t h e
minimum ceiling height for the
selection
mode.
Preferably
the
top of an outlet should be not
less than twice t h e o u t l e t ’ s
height below the minimum ceiling height.
8. Cfm/Sq
Ft Outlet Wall A r e a
is the stondord for judging total
room air movement. The maximum value shown results in on
air movement in the zone of occupancy of about 50 fpm. It is
orrumed
that furniture, pcopls,
etc.. obstruct 10% of the room
cross-vxtion.
If the obstructions
vary widely from
IO%, t h e
values
of the cfm/rq
ft outlet
wall
oreo
should be tempered
accordingly.
9 . F o r oppticotionr
requiring o
limiting
sound
level--the
outlet velocity is limited by the
round generated
by the outlet.
FACTOR
M a x Cfm/Sq F t
Outlet Wall A r e a
7.2
4.8
3.6
M i n Cfm/Sq F t
O u t l e t Wall A r e a
2.2
1.4
1.1
t
PIPING DESIGN
3-l
CHAPTER 1 m PIPING DESIGN-GENERAL
GENERAL SYSTEM DESIGN
Piping characteristics that are common to normal
air conditioning, heating and refrigeration systems
arc present4 in this chapter. The areas discussed
include piping material, service limitations, cspansion, vibration, fittings, valves, and prcssurc losses.
These areas are 01 prime consitlcration to the design
cnginecr since they influence the piping lile, maintenance cost and first cost.
The basic concepts ol fluid Ilow and design information on the more specialized fields such as high
temperature water or low tcmpcrature
refrigeration
systems are not included; this information is avnil;“.‘p in other authoritative sources.
MATERIALS
Tile materials most com~~lo~lly
systems arc the following:
I. Steel - black and galvani~ctl
used in piping
2. [\‘rought iron - black and galvanized
3. Copper - soft and hard
TtrOle I illustrates the recon~n~entled materials
[or the v a r i o u s scrviccs. Minimum standards, as
shown, should be maintainctl. Table 2 contains the
physical properties of steel pipe and Table 3 lists
the physical properties 0E copper tubing.
TABLE l-RECOMMENDED PIPE AND FITTING MATERIALS FOR VARIOUS SERVICES
SERVICE
Suction
.
REFRIGERANTS
12, 22, AND 500
Line
l i q u i d Line
Hot Gas Line
CHILLED
WATER
Wrought copper, wrought brass or tinned
cast brass
Steel pipe, standard wall
Lap welded or seamless for sizes larger
than 2 in. IPS
150 lb welding or threaded malleable iron
H&d copper tubing, Type L*
Wrought copper, wrought brass or tinned
cast brass
300 lb welding or threaded malleable iron
Hard copper tubing, Type L*
Wrought copper, wrought brass or tinned
cast brass
Steel pipe, standard wall
Lap welded or seamless for sires larger
than 2 in. IPS
~_ _~ .~__ .-~~ - --: : - -m
$%&or gglvantzed steel pipe@
L-.
--’
Galvanized
OR
WATER
DRAIN OR
CONDENSATE
LINES
STEAM OR
CONDENSATE
HOT WATER
-
-
’
Steel oiDe
Extia’strong wall for sizes 1 ‘/2 in. IPS and
smaller
Standard wall for sizes larger than 1 ‘/a
in. IPS
Lap welded or seamless for sizes larger
than 2 in. IPS
“Hbrd copp;;
CONDENSER
MAKE-UP
FITTINGS
PIPE
Hard copper tubing, Type L*
tub;&& -j’
steel
pipet
300 lb welding or threaded malleable iron
.%elding,- galvanizer.4;., c&t, malleab!ao
--“-&ckjron$ -4,.
-
r
9--.--^
_~
*Cast brass, wroughi &ape;pr y%wght brhsq
Welding,gaIvanized;castormalleableironf
Hard copper tubing+
Cast brass, wrought copper or wrought brass
Galvanized
Galvanized, drainage; cast or malleable
ironj
steel
pipet
Hard copper tubingi’
Cart brass, wrought copper or wrought brass
Black steel pipet
Welding or cast iron:
Hard copper tubingt
Cast brass, wrought copper or wrought brass
Block steel pipe
Welding or cast iron$
.Hard copper tubing+
‘2
.’
Cast brass, wrought copper or wrought brass
*Except for sizes l/4” and 3/g” OD where wall thicknesses of 0.30 and 0.32 in. are required. Soft copper refrigeration tubing may be used for sizes
1%” OD and smaller. Mechanical joints must not be used with soft copper tubing in sizes larger than %” OD.
tNormolly
standard wall steel pipe or Type M hard copper tubing is satisfactory for air conditioning applications. However, the piping material
selected should be checked for the design temperature-pressure ratings.
fNormally
125 lb cast iron and 150 lb malleable iron fittings are satisfactory for the usual air conditioning application. However, the fitting material
selected should be checked for the design temperature-pressure ratings.
3-1.
PAR'I- :\. I'II'II\:<; DESI(;N
\
TABLE 2-PHYSICAL PROPERTIES OF STEEL PIPE
INSIDE
DIAM
WEIGHT
OF
PIPE
(Ib/ft)
(in.)
.269
.215
10
12
14
16
,244
,314
,
,364
,302
,088
.I19
,424
,535
.0451
.0310
,141
,141
.0955
.0794
.1041
.0716
,493
.423
,091
,126
.567
,738
.oa27
13609
,177
,177
.1295
.I106
.I910
.I405
,622
.546
.109
.I47
.a50
1.087
.1316
.1013
,220
.220
.1637
.1433
.3040
.2340
.a24
,742
.113
,154
1.130
1.473
.2301
.I875
,275
,275
.216a
.194a
.5330
.4330
1.049
.957
.I33
,179
I.678
2.171
.3740
.3112
,344
.344
.2740
.2520
.a640
.7190
1.380
1.278
.I40
.191
2.272
2.996
.6471
.5553
.434
,434
.3620
.3356
1.495
1.283
1.610
1.500
,145
.200
2.717
3.631
.8820
.7648
,497
.497
.4213
3927
2.036
1.767
2.067
1.939
.154
,218
3.652
5.022
1.452
1.279
,622
,622
s401
JO74
3.355
2.953
2.469
2.323
.203
.276
5.79
7.66
2.072
i .a34
.7.53
,753
.6462
.6095
4.788
4.238
3.068
2.900
.216
.300
7.57
10.25
3.548
3.364
.226
.3ia
9.11
12.51
4.28
3.85
4.026
3.826
,237
.337
10.79
14.98
5.51
4.98
5.047
4.813
.25a
.375
14.62
20.78
6.065
5.761
.2ao
.432
i a.97
28.57
'12.51
11.29
7.981
7.625
,322
.500
28.55
43.39
21.6
19.8
2.24
2.26
WX)
80
10.750
10.750
10.750
10.020
9.750
9.564
,365
.500
.593
40.48
54.70
64.33
34.1
32.4
31.1
2.81
2.81
2.81
30(S)
40
00
a0
12.750
12.750
12.750
12.750
12.090
ii.938
11.750
11.376
.330
,406
,500
.687
43.80
53.53
65.40
88.51
49.6
48.5
46.9
44.0
306)
40
(Xl
80
14.000
14.000
14.000
14.000
13.250
13.125
13.000
12.500
.375
.43a
,500
.750
54.60
63.37
72.10
106.31
30(S)
40(X)
80
16.000
16.000
16.000
15.250
15.000
14.314
,375
,500
.a43
ia . 0 0 0
17.250
17.000
lb.874
16.126
.375
.500
.562
,937
40(S)
(5)
19.250
19.000
la.814
17.938
23.250
23.000
22.626
21.564
*TO change "Wt of Water in Pipe (lb/f1
.375
,500
,687
1.218
)
I
1.178
1.178
1.735
1.735
2.006
45.6
3.34
3.34
3.34
3.34
3.17
3.13
3.08
2.98
115.0
111.9
108.0
101.6
59.8
58.5
55.8
51.2
3.67
3.67
3.67
3.67
3.46
3.44
3.40
3.27
138.0
135.3
133.0
122.7
62.40
82.77
136.46
79.1
76.5
69.7
4.18
4.1 a
4.18
3.99
3.93
3.75
183.0
176.7
160.9
70.60
93.50
104.75
170.75
100.8
98.3
97.2
88.5
4.71
4.71
4.71
4.71
4.52
4.45
4.42
4.22
234.0
227.0
224.0
204.2
78.60
104.20
122.91
208.87
126.7
122.5
120.4
109.4
5.24
5.24
5.24
5.24
94.60
125.50
171.17
296.36
184.6
179.0
6.28
6.28
5.65
365.2
tz
1 6 ::
t o “ G a l l o n s o f W a t e r i n P i p e ( g o l / f t t r , "d k i d e d u e ti n t a b l e b y 8 . 3 4 .
t Si s d e s i g n a t i o n o f s t a n d a r d w a l l p i p e .
X i s d e s i g n a t i o n o f e x t r as t r o n g w a l lp i p e .
c
'
,
TABLE 3-PHYSICAL PROPERTIES OF COPPER TUBING
HARD
%%
Yi
I---
0 UTSIDE
IDIAM
STUBBS
GAGE
(in.)
OUTSIDE
SURFACE
(sq fl/fl)
,106
,144
,203
.036
,069
.I IO
,098
.I31
,164
1%
21
20
19
,032
.035
.042
.a1 1
1.055
1.291
,516
.a74
1.309
710
600
590
,328
,464
,681
,224
,379
.566
.229
.295
,360
I %
2
2%
18
17
16
.049
.058
,065
1.527
2.009
2.495
1.831
3.17
4.89
580
520
470
.94
1.46
2.03
,793
1.372
2.120
,425
,556
,687
15
14
13
,072
.oa3
,095
2.981
3.459
3.935
6.98
9.40
12.16
440
430
430
2.68
3.58
4.66
3.020
4.060
5.262
,818
,949
1.08
12
,109
,122
,170
4.907
5.881
7.785
18.91
27.16
47.6
400
375
375
6.66
8.91
16.46
8.180
11.750
20.60
1.34
1.60
2.13
19
,035
.040
,045
,430
,545
,785
.I46
.233
.484
1000
1000
1000
,198
,284
,454
,063
,101
.209
,131
,164
.229
.050
,055
,060
1.025
1.265
1.505
.a25
1.256
1.78
880
780
720
,653
,882
1.14
.35a
,554
,770
.295
,360
.425
1
Govt.Type
“K”
2 5 0 Lb
Working
PreSSWe
WT OF
WATER
INTUBE*
(lb/W
1000
1000
890
=/4
I-; vi
I
WEIGHT
OF
TUBE
(Ib/ft)
.083
,159
,254
A%%
SOFT
MINIMUM
TEST
PRESSURE
(psi)
,325
.450
,569
Govt.Type
"L"
2 5 0 Lb
Working
Pl.ZSSW.2
t
(in.)
TRANSVERSE
AREA
(rq i n . )
.025
.025
.028
1
G o v t . Type
“K”
4 0 0 Lb
Working
PreSSWe
INSIDE
DIAM
23
23
22
%
Gavt.Type
"M"
250 Lb
Working
WALL
THICKNESS
(in.)
1%
2
-
.070
1.985
3.094
640
1.75
1.338
.556
.,008900
,100
2.465 2.945
3.425
4.77 6.812
9.213
580 550
530
2 . 43 8. 3 3
4.29
2.070 2.975
4.000
.687 ,818
,949
.l10
.I25
.140
3.905
4.875
5.845
510
460
430
5.38
7.61
10.20
21
ia
I8
.032
.049
.049
.311
.402
,527
.076
.127
,218
1000
1000
1000
.133
,269
,344
,033
,055
.094
.098
.131
,164
16
16
16
.065
.065
.065
.745
.995
1.245
.436
.77a
1.217
1000
780
630
.641
.a39
1.04
.189
,336
.526
.229
.295
,360
15
14
13
.072
.oa3
.095
1.481
1.959
2.435
1.722
3.014
4.656
580
510
470
1.36
2.06
2.92
.745
1.300
2.015
,425
.556
.687
12
11
10
,109
.120
,134
2.907
3.385
3.857
6.637
a.999
il.68
450
430
420
4.00
5.12
6.51
2.870
3.890
5.05
8.18
,949
1.08
,160
.I92
4.805
5.741
18.13
25.88
400
400
9.67
13.87
11.97
18.67
26.83
5.180
8.090
11.610
7.80
11.20
1.08
1.34
1.60
1.34
1.60
%
f/
%
21
18
18
.032
.049
.049
.311
.402
.527
.076
.I27
.218
1000
1000
1000
,133
,269
,344
.033
.055
,094
.098
.131
.I64
% -
16
16
16
.065
.065
.065
.745
,995
1.245
,436
,778
1.217
1000
780
630
,641
.a39
1.04
,189
.336
.526
,229
,295
.360
15
14
13
.072
,083
.095
1.481
1.959
2.435
1.722
3.014
4.656
580
510
470
1.36
2.06
2.92
.745
1.300
2.015
,425
,556
,687
12
11
10
,109
.120
,134
2.907
3.385
3.857
6.637
a.999
11.68
450
430
420
4.00
5.12
6.51
2.870
3.89
5.05
.818
,949
1.08
.I60
.I92
4.805
5.741
18.13
25.88
400
400
9.67
13.87
* T o c h a n g e " W t o f w a t e r i n T u b e ( I b / f t ) " t o " G a l l o n s o f W a t e r i n T u b e ( g a i / f t ) , " d i v i d e v & e s i n t a b l eb y a . 3 4 .
7.80
11.2
1.34
1.60
,
PAR.l-
3-4
SERVICE
LIMITATIONS
3.
PII’Ih’(; I)I<SI(;N
TABLE J-THERMAL LINEAR EXPANSION OF
COPPER TUBING AND STEEL PIPE
The sal’c working pressure md teml~eratuw loI
steel pipe and copper tubing, including fittings, are
limited by the ASA codes. Check these codes when
thcrc is tiorrbt about the ability of pipe, tubing, or
fittings to withstand pressures and temperatures in a
given inst;~ll;ttion. In many instances cost can be reduced and over-design
ciiminatctl. For example, if
the working pressure is to bc 175 psi at 250 I;, a 150
psi, class A, carbon stcci flange can bc safely used
since it can withstand a pressure of 225 psi at 250 I;.
If the code is not checked, a 300 psi flange must be
spccifieti because the 175 psi working pressure exceeds the 150 psi rating of the 150 psi flange.
The safe working pressure and temperature for
copper tubing is dependent on the strength of the
fittings and tube, the composition of the solder used
for making a joint, and on the temperature of the
uid conveyed. Table 4 indicates recommended servIce limits for copper tubing.
(Inches per 100 feet)
TEMP
RANGE
COPPER
(F)
TUBING
STEEL
PIPE
0
50
100
0
.56
1.12
0
.37
.76
150
200
250
1.69
2.27
2.85
1.15
1.55
1.96
300
350
400
3.45
4.05
4.45
2.38
2.81
3.25
450
500
5.27
5.89
3.70
4.15
NOTE: Above data ore based on expansion from O’F but are rufficiently accurate for all other temperature ranges.
.
Chat -3 gives the sizes of offsets in steel piping
for travels up to 3 inches. Expansion loop sizes
may be reduced by cold-springing them into
place. The pipe lines are cut short at about
50:/, of the expansion travel and the expansion
loop is then sprung into place. Thus, the loops
are subject to only one-half the stress when expanded or contracted.
2. Expansion joints - There are two types available, the slip type and the bellows type. The
slip type expansion ,joint has several ciisadvantages: (a) It requires packing and lubrication,
which dictates that it be placed in an accessible
location: (b) Guides must be installed in the
lines to prevent the pipes from bending and
binding in the joint.
Bellows type expansion joints are very satisfactory for short travels, but must be guided or
EXPANSION OF PIPING
All pipe lines which are subject to changes in
temperature expand and contract. Where temperature changes are anticipated, piping members capable of absorbing the resultant movement must be
included in the design. Table 5 gives the thermal
linear expansion of copper tubing and steel pipe.
There are three methods commonly used to absorb
pipe expansion and contraction:
1. Expansion hops md 0Jfset.s - Table 6, page 6,
shows the copper expansion loop and offset
sizes required for expansion travels up to six
inches, Chn~t I shows the sizes of expansion
loops made of steel pipe and welding ells for
expansion travels up to 10 inches.
TABLE 4-RECOMMENDED MAXIMUM SERVICE PRESSURE FOR VARIOUS SOLDER JOINTS
MAXIMUM
SOLDER USED
IN JOINTS
50-50
Tin-lead
95-5
Tin-Antimony
95-5
Tin-lead
or
Solders Melting At
or Above
1100 F
SERVICE
VI
SERVICE
PRESSURE
(PSI)
Water
TEMP
100
150
200
250
vi”to
Incl.
200
150
100
85
100
150
200
250
500
400
300
200
350
270
1%”
I
l%“to2%”
IWI.
175
125
90
75
Steam
I
2vsfl
to 4%”
hl.
All
150
100
75
50
-
400
350
250
175
300
275
200
150
-
190
155
120
E x t r a c t e d f r o m A m e r i c a n Standard
W r o u g h t - C o p p e r a n d W r o u g h t - B r o n z e S o l d e r - J o i n t F i t t i n g s , (ASA 816.22-l 95 1
lisher, The American Society of Mechanical Engineers, 29 West 39th Street, New York 18, New York.
15
15
1, w i t h t h e p e r m i s s i o n o f t h e p u b -
C.HAP-I-ER
I, PIPISC; I)l-SI(;S
-
(;K:.Nf?RAI.
CHART 1 -STEEL EXPANSION LOOPS
AMOUNT
OF
WE V-0”
I
a
a
Q
Ib
12
,k
116
16
20
22
2l4
26
26
30
32
34
38
LENGTH OF “H”
Data from Ric-Wil Co.
CHART 2 -STEEL EXPANSION OFFSETS
LEG IN FEET
14
16
I6
20
22
24
26
28
30
32
34
36
38
40
LEG IN FEET
Data from Pittsburgh Pipe Coil L Bending Co.
TABLE 6-COPPER EXPANSION LOOPS
AND OFFSETS
EXPANSION LOOP
OFFSE?
01‘ s;~tltllcs slioultl be Llsetl. Tllc 1)i1)c s u p p o r t s illlist
Ii;kve a smooth, Ilat bearing surface, free from I)urrs
(jr otlicr sliarp projections which would wear or cut
tlic pipe.
7‘11~ controlling factor iI1 the spacing 01 supl)orts
LOI- horizontal pil)c lines is the ticllcction of piping
due to its own weight, weight of the. Iluiti, pi1)ing
accessories, and the insulation. Ttrble 7 lists tlic reconlnlcntlccl
support spacing for Sclictiulc -IO ljipe,
using tlic !istctl contiitioiis with watcl- as a fluid.
The support spacing for copper tubing is given
in TnDle 8 which inclutlcs the weight of the tubing
fillctl with water.
Data from Mueller Brass Co.
itI home other way restrained to prevent Callapse.
3. I;lexible ttzetd mtd rubber I~ose - F l e x i b l e
hose, to absorb expansion, is rcconmencied for
smaller size pipe or tubing only. It is not recommended for larger size pipe since an excessive length is required.
Where flexible hose is used to absorb expansion, it should be installed at right angles to
the motion of the pipe.
?‘hc devices listed above arc not always necessary
to absorb expansion and contraction of piping. In
fact they can be omitted in the great majority of
piping systems by taking advantage of the changes
in direction normally required in the layout. Consider, for example, a heat exchanger unit and a
primp located 50 ft. apart. Sufficient flexibility is
normally assured by running the piping up to the
ceiling at the pump and back down at the heat
exchanger, provided the piping is merely hung
lronl hangers and anchored only at the ends where
it is attachcti to the pump and the heat es&anger.
PIPING SUPPORTS AND ANCHORS
,111 piping shot~lti bc siipportcd with hangers that
cali withstand the combined weight of pipe, pipe
fittings, valves, iluitl in the pipe, and the insulation.
They must also IX capable of keeping the pipe in
proper alignment when necessary. IVlicre extreme
exlxinsion or contraction exists, roll-type hangers
Tables 7 clnd 8 arc for “ticad level” piping. \Vatcl
and refrigerant lines arc normally run level; steam
lines are pitched. Water lines are pitched when the.
line must be drained. For pitched steam pipes, rcfcr
t o Table 25, page 82, for support spacing when
Schedule 40 pipe is used.
Unless pipe lines are atlequatcly
anti properly
anchored, expansion may put excessive strain on
fittings and equipment. Anchors arc located according to jolt conditions. For instance, on a tall builcling, i.e. 20 stories, the risers could he anchored on the
5 t h Hoor and on the 15th Hoor with an expansion
device located at the 10th Hoor. This arrangement
allows the riser to expand in both directions from
the 5th and 15th floor, resulting in less pipe travel
at headers, whether they are located at the top or
bottom of the building or in both locations.
TABLE
7-RECOMMENDED SUPPORT
FOR SCHEDULE 40 PIPE
SPACING
NOMINAL PIPE SIZE
DISTANCE BETWEEN SUPPORTS
(in.)
(ft)
‘! - 1 %
8
1 vi - 2 %
3
- 3%
IO
12
4
14
_ 6
- 12
- 24
a
14
TABLE
16
20
8-RECOMMENDED SUPPORT
FOR COPPER TUBING
TUBE OD
(in.)
‘/I
vi
- I %
1 % - 2 ‘/r
2% - 5%
6% - 0 %
SPACING
DISTANCE BETWEEN SUPPORTS
(ft)
6
a
10
1 2
14
’
CHAPTER
011 smaller I~uildillgs, i . e . 5 s t o r i e s , riseI% arc
;I~j&(jre~l but O I I C C . Usually this is done near t h e
Ileader, allowing the riser to grow in one direction
only, either up or down depending on the header
location.
eI‘ll~ main point to consider when applying pipe
support anchors and expansion joints is that cxpansion takes place o n a temperature change. The
greater the temperature change, the greater the expansion. The supports, anchors and guides are applied to restrain the expansion in a desired direction
so that trouble does not develop because oE negligent design or installation. For example, if a takcoff connection from risers or heaclers is located close
to Iloors, beams or columns as shown in Fig.
a
change in temperature may cause a break in the
take 4 with subsequent loss of fluid and Hooding
da. Je. In. this figure trouble develops when the
riser expands greater than dimension “X.” Proper
consideration of these items is a must when designing piping systems.
VIBRATION
3-7
1. P I P I N G l>ESlGf% -GENERAL
ISOLATION
.OF
PIPING
SYSTEMS
The undesirable effects caused by vibration of
the piping are:
1. Physical damage to the piping, which results
in the rupture of joints. For refrigerant piping,
loss of refrigerant charge results.
2. Transmission of noise thru the piping itself
or thru the building construction where piping
comes into direct contact.
It is always difficult to anticipate trouble resulting
from vibration of the piping system. For this reason, recommendations made toward minimizing the
effects of vibration are divided into two categories:
’ design consideration - These involve design
i>recautions
that can prevent trouble effeclively.
FIG. 1 - T AKE -OFF Too CLOSE To FLOOK
2. liemedjes
or repai7.s - ‘i‘hese are necessary
where precautions arc not taken initially or,
in a minority of cases, where the precautions
prove to bc insuficicnt.
Design Considerations for Vibration Isolation
I. In all piping systems vibration has an originating source. This source is usually a moving
component such as a water pump or a compressor. When designing to eliminate vibration, the method of supporting these moving
components is the prime consideration. For
example:
a. The weight of the mass supporting the
components should be heavy enough to
minimize the intensity of the vibrations
transmitted to the piping and to the surrounding structure. The heavier the stabilizing mass, the smaller the intensity of the
vibration.
b. Vibration isolators can also be used to minimize the intensity of vibration.
c. A combination of both methods may be
required.
2. Piping must be laid out so that none of the
lines are subject to the push-pull action resulting from vibration. Push-pull vibration is best
dampened in torsion or bending action.
3. The piping must be supported securely at the
proper places. The supports should have a
relatively wide bearing surface to avoid a
swivel action and to prevent cutting action on
the pipe.
The support closest to the source of vibration
should be an isolation hanger and the succeeding hangers should have isolation sheaths as
illustrated in Fig. -7, p a g e 8. Non-isolated
hangers (straps or rods attached directly to the
pipe) should not be used on piping systems
with machinery having moving parts.
4. The piping must not touch any part of the
building when passing thru walls, Boors, or
furring. Sleeves which contain isolating material should be used wherever this is anticipated.
Isolation hangers should be used to suspend
the piping from walls and ceilings to prevent
transmission of vibration to the building.
Isolation hangers are also used where access to
piping is difficult after installation.
5. Flexible hose is often of value in absorbing
vibration on smaller sizes of pipe. To be effective, these flexible connectors are installed at
right angles to the direction of the vibration.
PART
3-8
3. PIPIN(; I)ESI(;K
c
DRAW UP SNUG
IL ,\ weight may IJC xtldecl t o tile pip bcfow
the first fixed support as illustrated in Fig. 3.
.l‘his weight adds mass to the pipe, reducing
vibration.
c. Ol,posing isolation hangers may be added.
FITTINGS
METAL
SLEEVE
Where the vibration is not limited to one
plane or direction, two flexible connectors arc
used and installed at right angles to each other.
The flexible hose must not restrain the vibrating machine to which it is attached. At the
opposite end of the hose or pair of hoses, a
rigid but isolated anchor is secured to the
connecting pipe to minimize vibration.
Generally, Hexible hose is not recommended in
systems subject to pressure conditions. Under
pressure they become stiff and transmit vibration in the same manner as a straight length
of pipe.
Flexible hose is not particularly efficient in
absorbing vibration on larger sizes of pipe.
EfFiciency i s i m p a i r e d s i n c e t h e length-todiameter ratio must be relatively large to obtain full flexibility from the hose. In practice
the length which can be used is often limited,
thus reducing its Hexibility.
Elbows are responsible for a large percentage of
the pressure drop in the piping system. With equal
velocities the magnitude of this pressure drop depends upon the sharpness of the turn. Long radius
rather than sflort radius elbows are recommended
wherever possible.
When laying out offsets, 45” ells are recommended
over 90” ells wherever possible. See Fig. 4.
Tees should be installed to prevent “bullheading”
as illustrated in Fig. 5. “Bullheading” causes tur- .
bulence which adds greatly to the pressure drop and
may also introduce hammering in the line. If more
than one tee is installed in the line, a straight
length of 10 pipe diameters between tees is recommended. This is done to reduce unnecessary turbulence.
To facilitate erection and servicing, unions and
Hanges are included in the piping system. They are
installed where equipment and piping accessories
must be removed for servicing.
The .various
methods of joining fittings to the
piping are described on pnge 12.
GENERAL PURPOSE VALVES
An important consideration in the design of the
piping system is the selection of valves that give
proper performance, long life and low maintenance.
Remedies or Repairs for Vibration Isolation After
Installation
1. Relocation OF the piping supports by trial and
error tends to dampen extraordinary pipe vibration. This relocation allows the piping to take
up the vibration in bending and helps to correct any vibrations which cause mechanical
resonance.
2. If relocation of the pipe supports does not
eliminate the noise problem caused by vibration, there are several possible recommendations:
a. The pipe may be isolated from the support
by nleans of cork, hair felt, or pipe insulation as shown in I;;#. 2.
I///////////
-FIXED
HANGER
WEIGHT
ADDED
CY
PUMP
FIG. 5 - WEIGHT
Ameo
TO
DAMWN VIBRATION
C:t-IAP-I’I:R
I. PIPIN<;
DI:SI(;N
-
3-9
(;ENERAL.
-T- -T-
- 4i
I
NO. I
“BULLHEADING”-
RECOMMENDED
NO. 2
PREFERABLE TO NO. I
ACCEPTABLE
*l-lit design, construction and niaterial of the valve
tlcternlines whether or not it is suited for the particalar application. Table 9 is for quick reference in
selecting a valve for ;I particular application. There
TABLE
9-GENERAL
NO.4
PREFERABLE TO NO. 3
PURPOSE
I
VALVES
STEAM
REFRIGERANT*
Satisfactory
Satisfactory
Satisfactory
Satisfactory
Satisfactory
Satisfactory (Low Press)
Satisfactory
Satisfactory
Satisfactory (High Press)
Satisfactory
Not Recommended
Not Recommended
Satisfactory
Satisfactory
Satisfactory
Satisfactory
Satisfactory
Satisfactory
(non-corrosive brines)
Satisfactory
Satisfactory
Satisfactory
Not Recommended
Satisfactory
Satisfactory
Not Recommended
Satisfactory
Not Recommended
Satisfactory
Satisfactory
Satisfactory
Satisfactory
Satisfactory
Satisfactory
(non-corrosive brines)
Satisfactory
Satisfactory
Satisfactory (Low Temp)
Satisfactory (Low Temp)
Satisfactory
Satisfactory
Not Recommended
R e c o m m e n d e d
Recommended
Satisfactory
Satisfactory
Satisfactory
Satisfactory
Satisfactory
Satisfactory
Satisfactory
Satisfactory
Satisfactory
Recommended
Not Recommended
Not
Not
Not
Not
G l o b e , A n g l e , “Y” V a l v e
I. P l u g d i s c
2. Conventional (Narrow-seat)
3. N e e d l e v a l v e
4. Composition disc
Satisfactory
Satisfactory
Satisfactory
Satisfactory
Satisfactory
Not Recommended
Satisfactory
Satisfactory (Low Press)
Satisfactory
Satisfactory
Satisfactory
Satisfactory
P l u g C o c k Valve
Satisfactory
Satisfactory
N o t R e c o m m e n d e d l’i
B. V a l v e s t e m , o p e r a t i o n
1. Rising stem, outside screw
2. Rising stem, inside screw
3. Non-rising stem, inside screw
4.
C.
Sliding
stem
Valve connections
1. Screwed
2. Welded
3. Brazed
4. Soldered
5. Flared
6. Flanged
to
Refrigerants
12,
22
and
...i
.i
’
pipe
DISC C O N S T R U C T I O N
Gate V a l v e
1. Solid wedge
2. Split wedge
3. Flexible wedge
4 . D o u b l e d i s c , p a r a l l e l seat
“For
- 47t
are six basic valves which are comnlonly used in
piping systems. These are gate, globe, check, angle,
“Y” and plug cock.
WATER
VALVE CONSTRUCTION
A. Bonnet and body connections
1. Threaded
2. U n i o n
3. Bolted
4. Welded
5. Pressure-seal
DO NOT USE
t
NO.3
PERMISSIBLE
500
only.
Recommended
Recommended
Recommended
Recommended
Each valve has a definite purpose in the control
of fluids in the system.
]~efore tliscussing the various valve types, the
construction dctails that arc similar in all of the
valves arc presented. These construction details are
made avaiIabIe to familiarize the engineer with the
various aspects of valve sclcction.
HANDWHEEL
(RISES WITH STEMI
,/-RISING
(,NSlDE
STEM
SCREW)
PACKING NUT
WITH GLAND
VALVE CONSTRUCTION DETAILS
XING BONNE
Bonnet and Body Connections
‘1%~ bonnet and body connections are normally
made in five different designs, namely threaded,
union, bolted, welded and pressure-seal. Each dcsign has its own particular use and advantage.
1. Threaded bonnets arc recommended for low
pressure service. They should not be used in a
piping system where there may be frequent
dismantling and reassembly of the valve, or
where vibration, shock, and other severe conditions may strain and distort the valve body.
Thrcadcd bonnets are economical and very
compact. Fig. 6 illustrates a threaded or
scrcwcd-in bonnet and body connection in an
angle valve.
2. U?rior~ I)ot~~~et and body construction is illustrated in I;ig. 7. This type of bonnet is normally not made in sizes above 2 in. because it
would require an extremely large wrench to
dismantle. A union bonnet connection makes
a sturdy, tight joint and is easily dismantled
and reassembled.
’ ’
RISING STEM
(INSIDE SCREW)
-
a
PLUG TYPE DISC
SCREWED
ENDS
.
FlC. ,i - GLCIUE
VALVE
3. Bolted bonnets are used on practically all large
size valves; they are also available for small
sizes. This type of joint is readily taken apart
or reassembled. The bolted bonnet is practical
for high working pressure and is of rugged,
sturdy construction. Fig. 8 is a gate valve illustrating a typical bolted bonnet a?d body
valve construction.
4. Welded bonnets are used on small size steel
valves only, and then usually for high pressure,
high temperature steam service (Fig. 9). Welded
bonnet construction is difficult to dismantle
x HANDWHEEL
ISE
(RISES WITH STEM1
-
PACKING NUT
WITH OUT GLAND
GLAND -
SCREWED
T H R E A D E D BONNET7
sot
ID WEDGE
D I S C .
‘BOLTED
BONNET
SCREWED
ENDS
NARROW SEAT DISC
FLOW
FLOW
FIG.
6 -ANGLE VALVE
FIG.~-GGATEVALVE(KISINCSTEM)
Figures 6.10, courtesy of Crane Co.
)
<:HAI’?‘EK
I. PII’INC; I)lCSl(;N
- (;ENERAL
3-U
Valve Stem Operation
In most applications the type ol stem operation
does not affect fluid control. However, stem construction may be important where the need for in.
dication of valve position is required or where head
room is critical. There arc four types of stem construction: rising stem with outside screw; rising
stem with inside screw; non-rising stem with inside
screw; and sliding stem (quick opening).
BOLTED
GLAND -
-WELDED
BONNET
FIG.
9
- W
ELDED
BONNET CONSTRUCTION
and reassemble. For this reason these valves
are not available in larger sizes. _
5. Aessure-seal bonnets are for high temperature
steam. Fig. 10 illustrates a pressure-seal bonnet
and body construction used on a gate valve.
Internal pressure keeps the bonnet joint tight.
This type of bonnet construction simplifies
“making” and “breaking” the bonnet joint in
large high pressure valves.
RISING STEM
(OUTSIDE SCREW)
HEEL
ES NOT RISE WrrH STEM)
‘f
FLEXIBLE WEDGE
DISC 7
FIG . 10 - FLEXIBLE
BOLTED GLAND
WELDING
W EDGE
DISC (GATE V ALVE)
1. Rising stem with outside so‘eru is shown in
Fig. 8. The gate valve illustrated in this figure
has the stem threads outside of the valve body
in both the open and closed position. Stem
threads are, therefore, not subject to corrosion,
erosion, sediment, and galling from extreme
temperature changes caused by elements in the
line fluid flow. However, since the valve stem
is outside the valve body, it is subject to damage when the valve is open. This type of stem
is well suited to steam and high temperature,
high pressure water service. A rising stem requires more headroom than a non-rising stem.
The position of the stem indicates the position (
of the valve disc. The stem can be easily lubricated since it is outside the valve body.
2. Rising stem with inside screw is probably the
most common type found in the smaller size
valves. This type of stem is illustrated in an
angle valve in Fig. 6, and in a globe valve in
Fig. 7. The stem turns and rises on threads
inside the valve body. The position of the stem
also indicates position of the valve disc. The
stem extends beyond the bonnet when the
valve is open and, therefore, requires more
headroom. In addition it is subject to damage.
3. Non-rising stem with inside screw is generally
used on gate valves. It is undesirable for use
with fluids that may corrode or erode the
threads since the stem is in the path of flow.
Fig. 11 shows a gate valve that has a non-rising
stem with the threads inside the valve body.
The non-rising stem feature makes the valve
ideally suited to applications where headroom
is limited. Also, the stem cannot be easily damaged. The valve disc position is not indicated
with this stem.
4. Sliding stern (quick opening) is useful where
quick opening and closing is desirable. A lever
and sliding stem is used which is suitable for
both manual or power operation as illustrated
in Fig. 12. The handwheel and stem threads
are eliminated.
3-12
L
,-NON-RISING
BOLTED
S
BONNET-
FLOW d
/~,,EwED ENDS
FIG. ll- GATEVALVE(NON-KISINGSTEM)
Pipe Ends and Vahe Connections
It is important to specify the proper end connection for valves and fittings. There are six standard
methods of joints available. These are screwed,
welded, brazed, soldered, flared, and flanged ends,
and are described in the following:
1.
ends are widely used and are suited
for all pressures. To remove screwed end valves
SLIDING STEM
CLAMP TYPE
BOLTED BONNET
FLANGED
.z / E N D S
A
FIG. If! -
SL.II)IM: hxar
GATE V ALVE
.
PAR.I- 3. I’IPINC;
I)ESI(;N
and fittings Irom the line, extra fittings (unions)
are required to avoid dismantling a consitlcrable portion of the piping. Screwed end connections arc normally confined to s~allcr pipe
silts since it is more difficult to make up the
screwed joint on large pipe sizes. I;ig. 7 is a
globe valve with screwed ends that connect to
pipe or other fittings.
2. Welded ends are available for steel pipe, fittings, and valves. They arc used widely for all
fitting connections, but for valves they are
used mainly for high temperature
and high
pressure services. They are also used where a
tight, leak-proof connection is required over
a long period of time. The welded ends are
available in two designs, butt weld or socket
weld. Butt weld valves and littings come in all
sizes; socket weld ends are usually limited to
the smaller size valves and fittings. Fig. 10
illustrates a gate valve with ends suitable for
welding.
3. Bl.ated ends are designed for brazing alloys.
This type of joint is similar to the solder joint
but can withstand higher temperature service
because of the higher melting point of the
brazing alloy. Brazing joints are use< principally on brass valves and fittings.
4. Soldered ends for valve and fitting are restricted to copper pipe and also for low pressure
service. The use of this type of joint for high
temperature service is limited decause
of the
low melting point of the solder.
5. Flared end connections for valves and fittings
are commonly used on metal and plastic tubing. This type of connection is limited to pipe
sizes LID to 2 in. Flared connections have the
advantage of being easily removed from the
piping system at any time.
6. Flanged ends are higher in first cost than any
of the other end connections. The installation
cost is also greater because companion flanges,
gaskets, nuts and bolts must be provided and
installed. Flanged end connections, although
made in small sizes, are generally used in
larger size piping because they are easy to
assemble and take down. It is v&y important
to have matching flange facing for valves and
fittings. Some of the common Range facings arc
plain face, raised face, male and female joint,
tongue and groove joint, and a ring joint.
Flange facings should never be mixed in.making up a joint. Fig. 8 illustrates a gate valve
with a flanged encl.
l
Cf-fAf’~I‘ER
I . PIf’IK(; I>ESI<;N
3-13
- (;I~NER.Al.
GATE VALVES
A gate valve is intcndcd for use iis a stop valve.
It gives the best service when used in the fully open
or closed position. f~i~7l7.e~ 8 (l?ld IO 111~21 14 arc typical gate valves con~n~only used in piping practice.
An important feature ol the gate valve is that
there is less obstruction and turbulence within
the valve and, therefore, a correspondingly lower
pressure drop than other valves. With the valve
wide open, the wedge or disc is lifted entirely out
of the fluid stream, thus providing a straight flow
area thru the valve.
I’
RISING STEM 2
( I N S I D E SCREW)
-- HANDWHEEL
-2l
IRISES WITH STEM)
----PACKING
NUT
WITH GLAND
SCREWED
BONNET
SPLIT WEDGE
DISC -
Disc Construction
Gate valves should not be used for throttling flow
except in an emergency. They are not designed lo1
this type of service and consequently it is difficult
to control flow with any degree of accuracy. Vibra1 and chattering of the disc occurs when the valve
1” Jsed for throttling. This results in damage to the
seating surface. Also, the lower edge of the disc may
be subject to severe wire drawing effects. The wedge
or disc in the gate valve is available in several
forms: solid wedge, split wedge, Hexible wedge, and
double disc parallel seat. These are described in the
following:
Solid wedge disc is the most common type. It
has a strong, simple design and only one part.
This type of disc is shown in Figs. 8 and 11. It
can be installed in any position without danger
of jamming or misalignment of parts. It is satisfactory for all types of service except where
the possi&lity of extreme temperature changes
exist. Under this condition it is subject to
sticking.
Split wedge disc is designed to prevent sticking, but it is subject to undesirable vibration
intensity. Fig. I3 is a typical illustration of this
type of disc.
Flexible wedge disc construction is illustrated
in Fig. 10. This type of disc is primarily used
for high temperature, high pressure applications and where extreme temperature changes
are likely to occur. It is solid in the center
portion ancl flexible around the outer edge.
This design helps to eliminate sticking and
permits the‘disc to open easily under all conditions.
LkmDle disc parallel seat (Fig. 14) has an internal wedge between parallel discs. Wedge
action damage at the seats is minimized and
transferred to the internal wedge where reasonable wear does not prevent tight closure.
FIG. 13 -SPLIT
Wmc;~ DISC
(GATE VALVE)
The parallel sliding motion of the discs tends
to clean the seating surfaces and prevents
foreign material from being wedged between :
disc and seat.
Since the discs are loosely supported except
when wedged closed, this design is subject to
vibration of the disc assembly parts when partially open.
(DOES NOT RISE
WITH STEM)
BOLTED
GLAND
RISING STEM
(OUTSIDE
SCREW
BOLTED BONNET -,
PARALLEL
DISCINTERNAL
WEDGEFLANGED
ENDS
FIG.
14 --DOUBLE DISC PARALLELSEAT
(GATEVALVE)
Figures
11.14,
courtesy of Crane Co.
I’.\l<-1‘ 3.
3-14
usctl in steam service, the closed valve
may trap steam between the discs where it contlenscs and creates a V~CLIUIII. This may result
ii1 leakage at the valve seats.
l’Il’IN(;
DESIGN
\Vhc~i
, ‘- HANDWHEEL
,RISES WlTH STEM,
,-PACKING
GLOBE, ANGLE AND “Y” VALVES
These three valves arc of the same basic design,
LIX and construction. They are primarily intended
Lor throttling service and give close regulation 0L’
flow. The method ol valve seating reduces
wire
drawing and seat erosion which is prevalent in gate
valves when used for throttling service
The angle or “Y” valve pattern is recommended
lor frdl flow service since it has a substantially
lower pressure drop at this condition than the globe
valve. Another advantage of the angle valve is that
it can be located LO replace an elbow, thus eliminating one fitting.
Z;ig. 7, pnge IO, is a typical illustration of a globe
valve, and Fig. 6, pnge 10, shows an angle valve. The
“Y” valve is illustrated in Fig. 15.
Globe, angle and “Y” valves can be opened or
closed substantially faster than a gate valve because
of the shorter lift oE the disc. When valves are to
be operated l’requently or continuously, the globe
valve provides the more convenient operation. The
seating surfaces of the globe, angle or “Y” valve
are less subject to wear and the discs and seats are
more easily replaced than on the gate valve.
Disc Construction
There are several different disc and seating arrangements for globe, angle and “Y” valves, each of
which has its own use and advantage. The different
types are plug disc, narrow seat (or conventional
disc), needle valve, and composition disc.
I. The p/zdg disc has a wide bearing surface on a
long, tapered disc and matching seat. This type
of construction offers the greatest resistance to
the cutting effects of dirt, scale and other
foreign matter. The plug type disc is ideal for
the toughest flow control service such as throttling, drip and drain lines, blow-off, and boiler
feed lines. It is available in a wide variety of
pressure temperature ranges. Fig. 7, pnge 10,
shows a plug disc seating arrangement in a
globe valve.
2. Nnrro~ sent (or conventional disc) is illustrated
in an angle valve in Fig. 6. This type of disc
does not resist wire drawing or erosion in
closely throttled high velocity flow. It is also
subject to erosion from hard particles. The
narrow seat disc design is not applicable for
close throttling.
NUT
, RISING STEM
, IINSIDE S C R E W ,
COMPOSITION DISC
SCREWED
BONNET\-_
SCREWED ENDS>
Courtesy of Jenkins Bras.
FIG. 15 - “Y”
VALVE
S. Needle v&es, sometimes referred to as expansion valves, are designed to give fine control of ’
flow in small diameter piping. The disc is
normally an integral part of the stem and has
a sharp point which fits into the reduced area
seat opening. I;&. 16 is an angle valve with
a needle type seating arrangement.
4. Composition disc is adaptable to many services
by simply changing the material of the disc.
It has the advantage of being able to seat tight
with less power than the metal type discs. It is
also less likely to be damaged by dirt or foreign
material than the metal disc. A composition
disc is suitable to all moderate pressure services
but not for close regulating and throttling.
Fig. 15 shows a composition disc in a “Y” valve.
This type of seating design is also illustrated
in I;&. 19 in a swing check valve and in Fig. 20
in a lift check valve.
RISING STEM
I INSIDE SCREW)
HANDWHEEL
( RISES WITH STEM
( w,THD”T GLAND)
NEEDLE DISC
FLUOW Y-
FK;. 16 -
SCREWED END
ANGLE V A L V E b\‘lTH r\iEEDLE
DI S C
CHAPTER 1. PlPlK(;
PLUG COCKS
I)ESI(;K
3-15
- (;ENEIIAL.
I
plug (.ocks are primarily used lor ixilancing in ;I
l)il)itlg system not subject to Ireclucnt changes in
110~~. They are normally less expensive than globe
tyl)e valves and the setting cannot be tampered with
as easily as a globe valve.
Plug cocks have approximately the same line loss
as ;I gate valve when in the fully open position.
\\%en lxirtially closed for balancing, this line loss
increases substantially. Z;ig. 17 is a lubricated type
plug valve.
HANDWHEEL
,RISES WITH STEM,
RISING STEM
lOUTSIDE SCREW,
-BOLTED BONNET
WIT” OIAPHRLGMSEAL,
FLANGED
ENDS
--
FLOW d
REFRIGERANT VALVES
Refrigerant valves are back-seating globe valves
of either the packed or diaphragm packless
type.
X’he packed valves are available with either a hancl
wheel or a wing type seal cap. The wing type seal
c“? is preferable since it provides the safety of an
. itional seal.
Where frequent operation of the valve is required, the diaphragm packless
type is used. The
diaphragm acts as a seal ant1 is illustrated in the
“E‘” valve construction in I;ig. 18. The refrigerant
valve is available in sizes up to IS/~ in. OD. For
larger sizes the seal cap type packed valve is used.
-’ COMPOSITION
FIG. 18 - "Y" V,\LVLC (DIAI~HKAGM I‘u15)
BOLTED
CHECK VALVES
‘COMPOSITION
There are two basic designs of check valves, the
swing check and the lift check.
The swing check valve may be used in a horizontal or a vertical line for upward flow. A typical
swing check valve is illustrated in Fig. 19. T h e f l o w
thru the swing check is in a straight line and without restriction at the seat. Swing checks are generally used in combination with gate valves.
FIG. 19 --SWING
CHECK VALVE
SCREWED UNION
RING BONNET
COMPOSITION DISC
FLOW
SCREWED END
FIG. 20 - LIFT
Courtesy
FIG. 15 -
of
Walworth Co.
PLUG C O C K
Ckik~K
VALVE
The lift check operates ill a manner similar to
that of a globe valve and, like the globe valve, its
llow is restricted as illustrated in Fig. 20. The disc
is seated by backflow or by gravity when there is no
Ilow, and is free to rise and fall, depending on the
pressure under it. The lift check should only be
installed in horizontal pipe lines and usually in
combination with globe, angle and “Y” valves.
Figuws II;, 18-20. courtesy of Crane Co.
VALVE AND FITTING PRESSURE LOSSES
SPECIAL SERVICE VALVES
‘J‘llc1.e ;tre scver;~i tyI)cs ol valves collllllorlly usctl
ii1 tliffcretit !)iping ;i!~plic;itions
t!l;It do IlOt nCccssarily fall i n t o the classificatiorl 01 general purpose
v;rlvcs.
b~xlxlnsion, relief, ant! solcnoit! valves are
COIIIC ol tlie more’ c o m m o n s!)eci;il purpose valves.
;\ rclicf valve is licltl closctl by ;I spring or some
otlicr means anti is tlesignet! to ;ititom;itically rclievc
tlic line or container pressure in excess of its setting.
III gcncr;il ;I relief valve should Ix installet! wherever
there is any danger of the fluid pressure rising above
tile tlcsign working pressure of the pi!x fittings
or !xessurc vcssc!s.
rJ‘o
properly design any type of piping system
a Iluir!, t h e losses thru t h e valves and
[ittings in clic system must IK realistically ev:~luatet!.
Tables hvc Ixxn prepared for cletermining these
losses in terms of ec!uivdcnt Icngth of pipe. These
values are then used with the correct friction chart
for the p;irticular fluid flowing thru the system.
TaOle /O gives valve losses with screwct!, Ilanget!,
llaret!, weltlcr!, soltlcret!, or Ixxzetl connections.
Tul~/e I/ gives fitting losses with scrcwet!, Ilan~ctl,
Il:irctl, welded, solclcretl, or brazed connections.
Tnble I2 lists the Iosses for special types of fittings
sometimes encounterctl in piping applications.
wnvcyin~
.
TABLE IO-VALVE LOSSES IN EQUIVALENT FEET OF PIPE*
Screwed, Welded, Flanged, and Flared Connections
GLOBE+
60° - Y
4s0 - Y
ANGLEt
GATEtt
i WING CHECKf
IF1 CHECK
NOMINAL
PIPE
.
OR
;;,,TUBE
SIZE
(in.)
.l%
J/
H
J/4
17
18
22
1%
29
38
43
15
20
24
12
15
18
2
2%
3
55
69
84
30
35
43
24
29
35
3%
4
5
100
120
140
6
8
170
220
280
1,
88
115
145
70
85
105
0.6
0.7
0.9
5
6
8
1 .o
1.5
1.8
10
14
16
24
29
35
2.3
2.8
3.2
20
25
30
41
47
58
4.0
4.5
6
35
40
50
70
85
105
12
14
16
360
410
130
155
180
18
20
24
460
520
610
200
235
265
*Losses
7
9
Globe &
Vertical
Lift
Same as
Globe
Valve**
60
80
100
Angle Lift
S a m e OS
Angle
19
22
25
165
200
240
VOlV.2
cue for all valves in fully open position.
tThese l o s s e s d o n o t a p p l y t o v a l v e s w i t h n e e d l e p o i n t t y p e s e a t s .
fLossa also a p p l y t o t h e i n - l i n e , b a l l t y p e c h e c k v a l v e .
**For “Y” pattern globe lift check valve with seat approximately equal to the nominal pipe diameter, use values of 60” “Y” valve for loss.
ttRegular and short pattern plug cock valves, when fully open, have some loss as gate valve. For valve losses of short pattern plug cocks
above 6 ins. check manufacturer.
clfnf’-r
r,.fc
I .
f’II’lN(.
Dk:SI(;I\‘--
(;E?Jl:RAI.
3-17
TABLE II-FITTING LOSSES IN EQUIVALENT FEET OF PIPE
Screwed,
Welded,
hOOTH
BEND
9o"
street*
NOMINAL
Flanged,
Flared,
and
Brazed
I
ELBOWS
45"
45"
Std*
street*
Connections
180’
Std*
SMOOTH
BEND
TEES
~
/Q
2.3
2.5
3.2
1
1%
1%
I
2.6
3.3
4.0
4
5
I
10
13
1.7
2.3
2.6
6.7
8.2
0.7
0.8
0.9
2.7
3.0
4.0
4.1
5.6
6.3
1.3
1.7
2.1
5.0
7.0
8.0
8.2
10
12
2.6
3.2
4.0
10
12
15
3.3
4.1
5.0
4.7
5.6
7.0
5.0
6.0
7.5
15
17
21
4.7
5.2
6.5
21
25
I8
5.9
6.7
8.2
8.0
9.0
12
9.0
10
13
25
-
12
I
14
16
30
34
38
18
20
24
42
50
60
I
*
0.9
1.0
1.4
1
1.7
2.3
2.6
-
16
18
20
68
78
19
23
26
29
33
40
-
23
26
30
85
100
115
29
33
40
60
MITRE
J/s
%
vi
1
1%
1%
I
1
ELBOWS
90° Eli
60' Eli
2.7
3.0
4.0
1.1
1.3
1.6
0.6
0.7
0.9
0.3
0.4
0.5
5.0
7.0
2.1
3.0
3.4
1.0
1.5
1.8
0.7
0.9
1.1
8.0
45'
EII
30'
EII
2
2%
3
10
12
15
4.5
5.2
6.4
2.3
2.8
3.2
1.3
1.7
2.0
3 %
4
5
18
21
25
7.3
0.5
11
4.0
4.5
6.0
2.4
2.7
3.2
6
8
30
40
50
13
17
21
7.0
9.0
12
4.0
5.1
7.2
10
/
2.3
3.1
3.7
1.4
1.6
2.0
1
2.6
3.3
4.0
7.9
10
13
19
23
26
NOMINAL
1.2
1.4
1.9
~jl
*R/D approximately equal to
1.
tR/D
approximately equal to 1.5.
1
/
26
30
35
40
44
50
I
1
30
34
38
42
50
60
'
P.\RT
3.
PIPING;
I)ESIGN
t
TABLE 12-SPECIAL
SUDDEN
‘h
ENLARGEMENT*
‘h
d/D
vi
FITTING LOSSES IN EQUIVALENT FEET OF PIPE
SUDDEN
%
CONTRACTION*
d/D
%
‘/a
SHARP
EDGE*
Entrance
PIPE
Exit
PROJECTION*
Enttmce
!
Exit
NOM.
18
20
24
-
-
-
-
(
18
-
-
20
-
-
-
*Enter table for losses at smallest diameter “d.”
/
1.5
1.8
2.8
.8
I .o
I .4
1.5
I .8
2.8
1.1
1.5
2.2
3.7
5.3
6.6
1.8
2.4
3.3
3.7
5.3
6.6
2.7
4.2
5.0
9.0
12
14
4.4
5.4
7.2
9.0
12
14
6.8
8.7
II
17
20
27
8.5
10
I4
17
20
27
13
I6
20
33
47
60
19
24
29
33
47
60
25
35
46
18
73
86
96
37
45
50
73
86
96
57
66
77
20
-
115
142
163
I
(
:1:
83
I
1
1::
163
!
1::
130
(
.
t
,
3-19
CHAPTER 2. WATER PIPING
Tliis chapter presents
the principles and currcntly accepted
design techniques [or water piping
systems i~s~tl in air conditioning applications. It also
includes the various piping arrangements for air
conditioning equipment and thestandard accessories
lountl in most water piping systems.
The principles and techniques described are
applicable to chilled water and hot water heating
systems. General piping principles and techniques
are described in Chapter I.
WATER
PIPING
SYSTEMS
Once-Thru and Recirculating
‘he water piping systems discussed here are di..:cl int’o once-thru and recirculating types. In a
dnce-thru system water passes thru the equipment
only once and is discharged. In a recirculating system water is not discharged, but flows in a repeating
circuit from the heat exchanger to the refrigeration
equipment and back to the heat exchanger.
Open and Closed
Both types are further classified as open or closed
systems. An open system is one in,which the water
flows into a reservoir open to the atmosphere; cooling towers and air washers are examples of reservoirs
oplen to the atmosphere. A closed system is one in
which the flow of water is not exposed to the atmosphere at any point. This system usually contains
an expansion tank that is open to the atmosphere
but the water area exposed is insignificant.
011 ijew co11struction.
The length of the water circuit
thru the supply ;incl return piping is tllc same for
aI1 uliits. Since the water circuits are equal for each
unit, the major advantage of a reverse return system
is that it seldom requires balancing. Fig. 21 is a
schclnatic sketch of this system with units piped
liori~ontally and vertically.
There are installations where it is both inconvenient and economically unsound to use a complete
reverse return water piping system. This somctimcs
exists in a building where the first floor has prcviously been air conditioned. To avoid disturbing the
lirst floor occupants, reverse return headers arc located at the top of the building and direct return
risers to the units are used. I;;!{,:. 22 illustrates a
reverse return header and direct return riser piping
system.
In this system the llow rate is not equal for all
units on a direct return riser. The difference in Ilow ’
rate depends on the design pressure drop of the
supply and return riser. This differcncc can be rctlucctl to practical limits. The pressure drop across
the riser includes the following: (I) the loss thru the
supply and return runouts from the riser to the unit,
(2) the loss tliru the unit itsell,
‘he recirculating system is further classified according to water return arrangements. When two
or more units are piped together, one of the following piping arrangements may be used:
1. Reverse return piping,
2. Kevcrse return header with direct return risers.
3. Direct return piping.
If the units have the same or nearly the same
Ill-essurc drop thru them, one of the reverse return
mcthotls of piping is recommendctl. However, if the
units have tliffercnt pressure drops or require balancing valves, then it is usually more ccononiical
lo use ;I direct return.
Reverse return piping is rcconinicntletl for most
c10Setl pipilig applications; it cannot be ~~sctl 011
(JpXl systems. It is often the most economical design
the loss thru
E
UNIT
Water Return Arrangements
:
and (3)
“NlTS
PIPED VERTICALLY
UNIT
SUPPLY
t
I
1
t
I
1
I
I
t
~RETURN
UNITS PIPED HORlZONTAl.LY
FIG. 2 1 - KEVE:RSE
KE-IIJRN P
I
I
I P I N G
I,
ii~21tlccl for a closed recirculating system where all
the units require Mancing valves ant1 have different i
prcssurc drops. Several Call-coil units piped tog&cl- ,
;iiid rccl~iiring tlilfercnt
w a t e r flow r a t e s , c;iIxicities *
and prcssurc (17-01)s i s 211 es;iil~~~le 01 t h i s type ol ’
\yalclll.
;
:
llle direct return piping system ‘is iriiYe&tly
u~ilx~la~icetl and rcquircs Ix~lancing valves or orifices;
and provisions to measure the pressirrc~drop in order
to meter the water flow. ~~\lthough Jnatcrial costs are
lower in this system than in the two reverse return
systems, engineering cost and balancing time often
otfsct this advantage. Fig. 23 illustrates units piped
vertically and horizontally to a direct return.
the fittings and valves. Excessive unlx~lancc in the
direct auI)IAy and return portion of the piping system ntay dictate the need for balancing valves or
orifices.
T o climinatc balancing valves, design the s u p p l y
and return pressure d r o p equal to one-fourth the
suni ol the pressure d r o p s 01 the preceding fteJJ7s
I, 2 (llld 3.
Direct retinxl piping is necessary for open piping
systems and is rcconimentlcd
for some closed piping
systems. A reverse return arrangement on an open
systcin requires piping that is normally unnecessary,
siiicc the same atmospheric conditions exist at all
opi points 01 the systcni. ,A direct return is rccom-
-E
SUPPLY
RETU6-4
UNITS
PIPED
I
I’
‘1
-P
UNIT
I’
PIPED
PIPING
DESIGN
‘I‘herc is a friction loss in any pipe thru which
water is flowing. This loss depends on the following
lactors:
I. Water velocity
2. Pipe diameter
:i. Interior surface
roughness
3. Pipe length
System pressure has no effect on the head loss ol
the equipment in the system. However, higher than
noriiial s y s t e m pressures may dictate the use of
heavier pipe, fittings and valves along with specially
tlcsignctl
cquipmcnt.
T O properly design a wat& piping system, the
cliginecr must evaluate not only the pipe friction
loss hut the loss thru valves, fittings and other equipJllent. In addition to these friction losses, the USC
of diversity in reducing the water quantity and
VERTICALLY
I
I
UNITS
WATER CONDITIONING
il:ormally all water piping systems must have aclequatc treatment to protect the various components
against corrosion, scale, slime and algae. Waier treatinent should always be under the supervision of a
water conditioning specialist. Periodic inspection of
the water is required to maintain suitable quality.
p~J1.l 5 of this manual contains a discussion Of the
various aspects of water conditioning including
cause, effect and remedies for corrosion, scale, slime
and algae.
WATER
UNIT
SUPPLY
CODES AND REGULATIONS ^.All applicable codes and regulations should be
checked to determine acceptable piping practice for
the particular application. Sometimes these codes
and regulations dictate piping design, limit the
pressure, or qualify the selection of materials and
equipment.
HORIZONTALLY
.
l
CHAPTER 2. \V.\TER
3-21
l’II’IK(;
TABLE 1 I-RECOMMENDED WATER VELOCITY
VELOCITY
SERVICE
PIPE FRICTION LOSS
‘1‘11~ l)il)c [rictioll loss
water
1)~s~
\,clo(:ity,
iI1
:I
systelll
pipe diameter, interior
;lll(l piljc lellgth. Varying
tk~Kl~&
slll~Llc:c
(fps)
011
rough-
any ollc OL t h e s e hc-
tars inlluences the total friction loss in the pipe.
,\lost air conditioning ;il)piications use either steel
l)ipe (jr (:oppu- tubing in the piping system. 7’0 CKlIll;ltC: the [riction loss in steel pipe or copper tubing,
refer to Cl~0j’t.s j 111~71 5 in this chapter.
Clrtl72.y j and f are for Schedule 40 pipe up to 2J
in. in diameter. CAn7.t j shows the friction losses for
closed recirculation piping systems. The friction
losses in C/l& f are for open once-thru and for open
reci- .!iation piping systems.
( ,,.,r~t j shbws friction losses for Types K, L and RI1
copper tubing when used in either open or closed
water systems.
These charts show water velocity, pipe or tube
diameter, and water quantity, in addition to the
friction rate per 100 ft of equivalent pipe length.
Knowing any two of these factors, the other two can
be easily determined from the chart. The effect of
inside roughness of the pipe or tube is considered
in a11 these values.
The water quantity is determined from the air
conditioning load and the water velocity by prcdetermined recommendations. These two factors arc
used to establish pipe size and friction rate.
Water
RANGE
Velocity
The velocities recommended for water piping
tlepencl on two conditions:
The service for which the pipe is to be used.
2. The effects of erosion.
Table 13 lists recommended velocity ranges for.
different services. The design of the water piping
system is limited by the maximum permissible flow
velocity. The maximum values listed in Table 13
are based on established permissible sound levels of
moving water and entrained air, and on the effects
of erosion.
Erosion in water piping systems is the impingement on the inside surface of tube or pipe of rapidly
nioving water containing air bubbles, sand or other
solid matter. In som&cases
this may mean complete
deterioration of the ‘tube or pipe walls, particularly
on the bottom surface and at the elbows.
Since erosion is a function of time, water velocity,
and suspended materials in the water, the selection
of a design water velocity is a matter of judgment.
‘1%~ ni;ixiinuIIL water velocities presented in ?‘rlOle
If
;lix based on many years
of
cxpericncc
2nd t h e y
insure the attainment of optimum equipment lilt
unticr normal conditions.
Friction Rate
The design of a water piping system is limited by
the friction loss. Systems using city water must have
the piping si/.etl so as to provide the required BOW
rate at a pressure loss within the pressure available
at the city main. This pressure or friction loss is to
include all losses in.the system, as condenser pressure
,tirop,. pipe and fitting losses, static head, and water
meter drop. The total system,pressure drop must be
less than the city main pressure to have design
water flow.
A recirculating system is sized to provide a rcasonable balance between increased pumping horsepower
due to high friction loss and increased piping first
cost due to large pipe sizes. In large air conditioning
applications this balance point is often taken as a
maximum friction rate of 10 ft of water per 100 ft
of equivalent pipe length.
In the average air conditioning application the
installed cost of the water piping exceeds the cost
of the water pumps and motors. The cost of increasing the pipe size of small pipe to reduce the friction
rate is normally not too great, whereas the installed
cost increases rapidly when the size of large pipe
(approximately 4 in. and larger) is increased. Smaller
pipes can be economically sized at lower friction
rates (increasing the pipe size) than the larger pipes.
In most applications economic considerations dictate that larger pipe be sized for higher flow rates
TABLE 14-MAXIMUM WATER VELOCITY
TO MlNlMlZE EROSION
NORMAL
OPERATION
(hr/vd
1500
2000
3000
4000
6000
8000
WATER,VELOCITY
(fPf)
12
11.5
11
10
9
8
CHART 3-FRICTION
LOSS FOR CLOSED PIPING SYSTEMS
Schedule 40 Pipe
.I
15
\
20000
15000
2.0 .25
.3
.4
.5
.6
.8 1.0
1.5
\
\
\
10000
e 000
6 0 0 0
5000
4 0 0 0
3000
2000
.
I500
IO00
800
\
600
j,
600
500
400
3 0 0
200
150
100
80
6 0
60
50
5 0
4 0
4 0
3 0
2 0
15
IO
8
6
6
5
5
4
4
FRICTION LOSS (FEET OF WATER PER 100
FT)
3-23
CHAI’TER 2. WATI~K PIPING
CHART 4-FRICTION LOSS FOR OPEN PIPING SYSTEMS
Schedule 40 Pipe
I
.I5
.2
.25
.3
4
5
.6
.0
II
1.0
1.5
2 . 0 2.5 3
4
5
6
0
IO
I5
2 0 25
30
4 0
60
00
100
N
10000
0 0 0 0
6 0 0 0
6 0 0 0
5000
5 0 0 0
4000
4 0 0 0
3000
2000
I500
IO00
000
600
500
300
z
a 2 0 0
cl
3
s
L
150
100
100
00
00
60
6 0
50
5 0
4 0
4 0
3 0
20
I5
IO
IO
0
7
0
7
6
6
5
5
4
4
3
2
1.5
1.0
FRICTION LOSS (FEET OF WATER PER IOOFT.)
L
3-24
CHART S-FRICTION LOSS FOR CLOSED AND
Copper
4000
I
.I5
.2
.25 .3
.4
.5
.0
.6
I.0
1.5
2
2.5
I’,\R’I’
3.
I’II’INC lIESIGN
OPEN PIPING SYSTEMS
Tubing
3
4
010
56
15
20
25
30
40
50
60
3000
2000
I
ITYPk
Id
U+-I-I+l
i
1500
1000
000
000
600
6 0 0
500
500
400
4 0 0
300
3
4 0
s
L
3 0
2 0
15
6
5
4
.I
.I5
.2
.25 .3
.4
.5
.6
.0
1.0
FRICTION
1.5
LOSS
2
2.5
(FEET
3
OF
4
5
6
W A TER
0
IO
15
PER loons.)
20
25 30
4 0
50 60
00
1.0
100
’
;IIJ([ l)r(‘ssurc drops that~ smaller l)ipc wliic.11 is si/ctl
1’0~ l~mu- prcssurc drops ant1 (low rates.
l&cptiolls to tllis general guide oltcn occur. I;or
appearance or physical limitations may
tli(,t;lte the use 01 sriiall pil)cs. Tllis is ol’ten done I’or
short runs where the total 1)ressurc tlrol) i s n o t
greatly inllucnced.
E;~ch s y s t e m al~oultl be alialy~ccl scparatcly t o
tleternline tlic economic balance between first cost
(pipe si/e, pump ;intl motor) and operating cost
(1)rcssure drop, pump and motor).
esmple,
Pipe Length
-1‘0 clctcrminc the friction loss in a water piping
straight
system, the engineer must calculate the
lengths ol pipe and evaluate the additional cquivalent lengths of pipe due to fittings, valves and other
‘~lents in the piping system. Tn1)le.s IO, If and 12
b me the ‘additional equivalent lengths of pipe for
these various components. The straight length of
pipe is measured to the centerline of all fittings and
valves (Fig. 2-f). The equivalent length of the components must be added to this straight length oE
pipe.
WATER PIPING DIVERSITY *
LVhen the air conditi6ning load is determined for
each exposure of a building, it isassumed that the
exposure is at peak load. Since the sun load is at
a maximum on one exposure at a time, not all of
the units on all the exposures require maximum
water How at the same time to handle the cooling
load. Units on the same exposure normally require
maximum How at the same time; units on the
adjoining or opposite exposures do not. Therefore,
if the individual units are nzrtonznticnlly
controlled
rary
the
water
quantity,.the
system
water
quantity
‘\
sicually required during normal operation is less
than the total water quantity required for the peak
design conditions for all the exposures. Good engineering design dictates that the water piping and
the pump be sized for this reduced water quantity.
The principle of diversity allows the engineer to
evaluate and calculate the reduced water quantity.
In all water piping systems two conditions must be
satisfied before diversity can be applied:
1. The water How to the units must be automatically controlled to compensate for varying
loads.
2. Diversity may only be applied to piping that
supplies units on more than one exposure.
I;iRlye 25 is a typical illustration of a header layout to which diversity may be applied. In this il-
FK.
24-
PIPL
LENGTH MEASUREMENT
lustration the header piping supplies all I‘our
exposures. Assuming that the units supplied are
automatically controlled, diversity is applied to the ’
T\‘est, south and east exposures only. The last leg or
exposure is never rcducecl in water quantity or pipe
size since it requires full flow at some time during
operation to meet design conditions.
Fi,qxye 16 illustrates another layout where diversity may be used to reduce pipe size and pump
capacity. In this illustration diversity may be applied
to the vertical supply and return headers and also to
the supply and return branch headers at each Boor.
Diversity is not applied to pipe section ‘i-8 of both
the supply and return vertical headers. In addition
N
W
FK. 25 - HEAL)EK
I’II’INC
3-26
I’/\Rl’
3. I’II’INC;
I>ESI(;N
4
Solution:
I . l’utnp .\ srtpplics nor111 ant1 west cxl)~~s\Irc
I)llt tlivcrsity
can IX applic~l tn norrh cxpos”rc only. The tolal g~“n in
p~~n,p ;\ circuit is 280 gpm and the a~~~mulatctl gpn’
in the north exposure is i(iO gpm. The ratio of accllmlliatctl gpm to the total waler quantity in the circuit is:
I GO
280
Enter Clrccrl
factor ,785.
Pump
..)I
6 at the ratio 57 and rcatl tllc t l i v c r s i t y
1% circuit has a ratio for the cast exposure of:
120
"80=.43
Entering Chart 6 at the
is read as ,725.
2.
ratio of .43,
the
diversity factor
The following table illustrates how the diversity factors
are applied to the maximum water quantities to oljtain
the design water quantities.
.
I’UMI’ “A” CIRCUI-I
Max
Quantity
(mm)
Section
F IG.
26 - HORIZONTAL \VVA.TER
PMNG
LAYOUT
the south leg of the return piping and the west leg
of the supply piping on each floor must be full size.
In any water piping system with automatically
controlled units, the water requirements and pump
head pressure varies. This is true whether or not
diversity is applied. However the water requirements
and pump head vary considerably more in a system
in which diversity is not considered.
In a system in which diversity is not applied,
greater emphasis is required for pump controls to
,’
revent excessive noise being created by throttling
%-. alves or excessive w5ter velocities. In addition, since
the system never requires the full water quantity
for which it is designed, the pump must be either
throttled continuously, or bypassed, or reduced in
size.
It is good practice, therefore, to take advantage
of diversity to reduce the pipe size and pump capacity. Chn~t 6 gives the diversity factors which arc
used in water piping design. Exnmple I illustrates
the use of Chn~t 6.
Example 1
- Diversity Factors
for Wafer Piping
Given:
Water piping layout as illustrated in Fig. -77.
Find:
I. Diversity factor to I)e applied to the water quantity.
2. Water quantities in header sections.
Headers
Design
Quantity
(gpm)
Diversity
Factor
A-RI
RI-R2
R2-R3
R3-R4
R4-R5
280
260
240
220
200
.785
.785
.785
.785
,785
220
204
188
I73
157
R.‘,-RG
R6-R7
R7-R8
KS-R%
R9-RIO
180
160
140
I20
100
,785
,785
,785
I41
* 126
RIO-RI1
Rll-RI2
Rl2-RI3
Rl3-RI4
I
1
80
60
40
20
I
I
P U M P “B”
I .oo
1 .oo
1.00
1 .oo
I .oo
I .oo
I
(110)120*
120
100
80
GO
40
20
CIRCUIT
B-R28
R28-R27
R27-R26
R2F-R25
R25R24
R24-R23
R23-R22
R22-R2l
R2l-R20
R20-RI9
280
260
240
220
200
160
140
120
100
I
1
I
I
.oo
.oo
.oo
.oo
160
140
120
100
Rl9-R18
Rl8-RI7
R17-RI6
RI&RI9
80
. 1.00
I .oo
I .oo
I .oo
80
180
GO
40
20
,725
.725
,725
,725
.725
203
188
174
( 1 4 5 )IGo*
.725
( 1 3 0 )lGO*
160
GO
40
20
*When applying diversity, the design water quantity in
the last section of the exposure is usually less than the
water quantity in the first section on the adjoining exposure. When this occurs, the water quantity in the last
section or last two sections is increased to equal the
water quantity in the first section of the next exposure.
CH.\ P~I‘Icl<
2. wA’I‘I:Iz
3-27
PI 1’1 xc;
CHART
.I0
.20
.30
S--DIVERSITY
.40
50
FACTORS
.60
.70
.80
.90
.I00
ACCUMULATED WATER FLOW FOR EXPOSURE
TOTAL WATER FLOW FOR PUMP
N
20
In Example 1 pump “A” is sclcctcd for 220 gpm
ant1 pump “B” is selected for 203 gpm. The pipe skes
in the north and east exposures are reduced using
the design gpm, whereas the pipes in the south
and west exposures are select4 lull size.
Esrrttqle -3 077d 3 illustrate the economics involved
when applying diversity. Esatttple 2 shows ;I typial
header layout with one 1)ump serving ~11 (our esposures. The header is sized without diversity.
Exrrtttple 3 is the sanic piping layout Ijut
is used to size the heatlcr.
tlivcrsity
2.
;I
20-m
RIO
ZO-
RII
20-
RI2
20-
RI3
2d
RI4
RI5
20
20
20
20
20
A7
R6
R5
R4
R3
‘”
2 0 p.
PLAN
VIEW
W
20
20
E
RI6
RI7
RI8
RI9
R20
20
20
20
20
20
S
R23
R2l R 2 2
20
\
20
20
Eli,ows. R/I) - 1
Kxpcctcd Icngth of operation - 6000 Itours
NORTH
RI
R2
\
R4
R3
U
U
A
R5
u
U
320
3 4 0 1
7.0’
I
40
20’
760
7
FIG.
Solution:
I. Design water velocity for sizing the hcatlers is determined
from Tnl~les 13
14.
Slaximum water velocity - i fps
2 . \laximum water quantity required when no diversity
is applied is 3G0 gpm. Pump is selected for 360 gpm.
c
240
PLAN VIEW
20’
;;;I
60 4’1
I:intl:
I. lksign header water velocity
2. \Vater quantity for pump selection
3. Header pipe size and pump friction head
I-
20’
100
n
,
1
20’
120
rl
(
20’
1
140
r-l
,
3. The table below gives the header pipe sizes and pump
Friction head when no diversity is applied.
20’
1
I60
160
n
Example 3 - Sizing Header Using Diversity
28 - SUPPLY HEADER PIW SIZING
Example 2 - Sizing Header Using No Diversity
Given:
.\ building with a closed recirculation water piping system
using a horizontal header and vertical risers as illustrated
in Frrg-. 28.
1laximum llow to each riser - 20 gpm
Schetlulc 40 pipe and fittings.
Given:
Same piping layout as in Exa+e -7 and Fig. 28.
Maximum flow to each riser - 20 gpm
Schedule 40 pipe and fittings
Elbows, R/D = 1
Expected length of operation - 6000 hours
Maximum design velocity - 7 fps (Exantple 2)
Find:
I.
2.
3.
4.
.
Diversity factor for each exposure
Design gpm for each header section
JVater quantity for pump selection
Header pipe size and pump friction head
.
H E.-\DER
SECTION
TORI
Rl-R2
R2-R3
R3-R4
R4-R5
PIPE
SIzEt
kP’)
(in.)
SF0
340
320
300
280
-
5
5
5
.i
4
’
.
LENGTF
RETWEE:
T.iKEOFFS
w
-
27
I8
20
20
20
260
41
RF-R;
R7-R8
R8-R9
240
220
200
4
-I
4
20
20
20
R9-RIO
180
4
8
RIO-RI1
Rll-RI2
R12-RI3
160
140
120
3
3
3
20
20
20
R13-RI4
RI4R15
100
80
RlS-RI6
R16-R17
R17-RI8
GO
-10
2 0
R5-R6
‘-
W.\TER
OUAN?I-ITY
8
FITTIXGS
-
FITTING’
EQUIVALENT
LENGTH*
w
TOTAL
EQUIV.-\LENT
LENGTH
(4
FRICTION
LOSS?
(ft of water
per 1 0 0
equiv ft)
FRICTION
HE.AD
(ft of water)
2-ells .I-tee
l-tee
l-tee
1 -red. tee
26
8.2
8.2
8.2
12.0
53.0
26.2
28.2
28.2
32.0
2.3
2.0
I .8
16
4.4
1.22
.53
.dl
.45
I.41
l-tee
I-ell
I-tee
l-tee
l-tee
6.7
10.0
6.7
6.7
6.7
24.7
26.7
26.7
26.7
3.8
3.2
2.7
2.3
.94
l-tee
1 -ell
l:red. tee
l-tee
l-tee
6.7
6.0
9.0
5.0
5.0
20.7
29.0
25.0
25.0
2.1
5.5
4.6
3.2
.43
I .59
I.15
.80
l-tee
l-tee
1-ell
1 -red. tee
1 -tee
1 -red. tee
5.0
!‘,.O
7.5
7.0
3.3
5.0
25.0
2.5
20.5
27.0
23.3
25.0
I.6
6.8
3.2
B.5
Pump
friction
.85
.72
.62
.33
I .84
.i5
I .62
headf 16.39
*Fitting losses are determined from Tnhlr II. For reducing tees enter Table I! at the larger diameter.
tFriction
rate and pipe size are determined from Cltnrr 3 not exceeding the maximum design water velocity (7 fps).
:Pump friction head does not include losses For valves. strainers, etc., which must he included in the actual design.
(:HAPTER
2 .
WA’I‘ER
3-29
I’II’INC
DINtiN
I . Cl~ur-t 0 i s risccl with tllc ratio of accumulated
gpm i n
rhc exposure to the total pump gpm, in order to dctcrfollowing talk illustrates
mint l h c diversity factors. The
the method of determining diversity factors. (First cxposure listed is always first exposure served by pump.)
l<SI’OSI!RI:
North
East
S011th
ivcst
=$yJ;E
ACCtihI.
Gl’hl
DIVERSITY
FACTOR
TOTAL
I’UXII’ Gl’hf
Q”,;NTrTY
(gP’n)
100
80
100
HO
100/360 = .28
180/360 = .5O
.Gi
.5G
280/360= .i8
3 6 0 / 3 G O= I . o o
239
I .oo
2. The diversity factor found in SteL I is applied to the
maximum water quantity in each header section to
cstahlish the design gpm for sizing the header. The table
at right gives the design water quantity for the various
header sections.
3.
4.
The design water quantity required
when diversity is applied is 240 gpm.
for pump
HEADER
SECTIOS
4
4
4
4
4
27
18
20
20
-20
115.RG
197
4
8
RF-R7
R7-R8
RS-R9
182
4
167
160
4
3
20
20
20
R9-RIO
160
3
8
Rl3-RI4
R14-RI5
90
R15-RI6
Rl6-RI7
Rli-R18
GO
80
40
20
20
20
20
20
8
20
20
20
260
240
220
200
.7G
7G
.7G
.iG
1
160
140
120
100
39
39
142
.89
.89
12-i
1 .oo
106
90
80
1.00
I .oo
1 .oo
GO
40
20
,
GO
/
40
20
(lH7)197#
I97
182
lG7
(152)160f
160
.89
80
/
Rl5-RIG (
Rl6-Rl7
R17-R18
*lVhen applying diversity, the design water quantity in
the last section of the exposure is usually less than the
design water quantity in the first section of the adjoining
exposure. When this occurs, the water quantity in the
last section or last two sections is increased to equal the
water quantity in the first section of the next exposure:
(in.)
240
227
ii7
67
.G7
Ai7
IL
RlO-RI 1
Rl l-R12
Rl2-RI3
Rl3-RI4
R14-RI,5
PIPE
SIZEt
To Rl
RI-R2
R2-R3
R3-R4
R4-R3
2 14
201
197
R5-RB
RG-R7
R7-RH
R&J-R9
180R9-RlO
340
320
300
280
selection
‘rhe design water quantity found in Strp -3 is used in
sbing the header pipe and in establishing the pump
friction head. The talk Ijelow illustrates the header pipe
sizing:
DESIGK
W’.\TER
QUANTITY
(gpm)
RI-R2
R?-KS
RS-R4
R4-R5
\V.\‘I‘ER
(,)LJ.\N’[‘I’I‘Y
kI”“)
240
““7
..h
2 I4
201
2-ells
l-tee
l-tee
1 -tee
l-tee
20.0
6.7
6.7
6.5
6.7
l-e11
l-tee
l-tee
l-tee
1 -red. tee
10.0
6.7
6.7
6.7
9.0
1 -ell
l-tee
l-tee
1-tee
l-tee
7.5
5.0
5.0
5.0
5.0
l-tee
1-ell
l-tee
l-red. tee
l-tee
1 -red. tee
5.0
7.5
4.0
7.0
3.3
5.0
i
!I
TOTAL
EQUIVALENT
LENGTH
(W
FRICTION
Losst
(ft of water
per 100
equiv ft)
FRICTION
HEAD
47.0
24.7
26.7
26.7
26.7
3.4
3.0
2.7
2.3
2.3
1.60
24.7
26.7
26.7
29.0
2.3
2.0
1.8
5.6
57
20.5
25.Q
25.P
25.0
5.6
4.5
3.5
2.7
(ft of water)
.74
.72
.61
.61
.53
.48
1.62
1.12..
.8i
68
25.0
20.5
27.0
23.3
25.0
Pump
friction head: 16.35
‘Fitting losses are determined from Tnble 11. For reducing tee enter Table II at the larger diameter.
tFriction rate and pipe size are determined from Chart 3 with the water velocity not exceeding 7 fps.
:Prlmp friction head does not include losses for valves, strainers, etc., which must be included in the actual
design.
3l30
i
lhcitrtp1e.r
2 clrlfl 3 iridicalc that the following
reductions iI1 pipe ant! fitting siLc c;lIl be made when
diversity is used:
I. 5 j ft of 5 in. pipe replaced with 4 in. pipe.
2. 28 ft of ‘1 in. pipe replaced with 3 in. pipe.
3. 8 fittings rccluced 1 size.
In addition the pump can be selected for 240 gpm
instead of 360 gpm which is approximately a l/s
reduction. Other areas whcrc a reduction in size is
possible are:
1. Pipe and fitting5 in the return piping header.
2. Valves, unions, couplings, strainers and other
clcmcnts located in the supply and return
licadcrs.
PUMP SELECTION
Pumps are selected so that there is no sustained
ise in pressure when the water Row is throttled.
Systems having considerable throttling have the
pump selected on the flat portion of the “headversus-flow” curve.
Normally, new installed pipe has less than design
friction and, therefore, the pump delivers greater
gpm than clesign and requires more horsepower.
For this reason a centrifugal pump is always selected
for the calculated pump head without the addition
of safety factors. If the pump is selected for the
calculated head plus safety factors, the pump must
handle a larger water quantity. When this occurs
and provision is not made to throttle or bypass the
excess water flow, the possibility of pump motor
overload exists.
Again, if the pump is selected for maximum
water quantity and diversity is not applied, the
water flow must be throttled. This increases the
->ump head.
SYSTEM ACCESSORIES AND LAYOUT
I’i\K’I‘ 3. PIPING DESIGN
The open and closed expansion tanks are the
two types used in water piping systems. Open expansion tanks are open to the atmosphere and are
located on the suction side of the pump above the
highest unit in the system. At this location the tank
provides atmospheric pressure at or above the pump
suction, thus preventing air leakage into the system.
‘I’he static head on the pump due to the expansion
tank must be greater than the friction drop oE the
water in the pipe from the expansion line connection to the pump suction. In Pig. 139 the static
head AB must be greater than the friction loss in
line AC. Adding any accessories such as a strainer
in line AC increases the friction drop in AC and
results in raising the height of the expansion tank.
‘To keep the height of the tank at a reasonable level,
accessories should be placed at points 1 and 2 in .
I;is.
< 29. At these designations the friction loss in
line AC is not affected.
The following procedure may be used to determine the capacity of an open expansion tank:
1. Calculate the volume of water in the piping,
from Tables 2 and 3, pages 2 and 3.
2. Calculate the volume of water in the coils and
heat exchangers.
3. Determine the percent increase in the volume
of water due to operating at increased temperatures from Table 15.
4. Expansion tank capacity is equal to the percent
increase from Table 15 times the total volume
of water in the system.
The closed expansion tank is used for small or
residential hot water heating systems and for high
temperature water systems. Closed expansion tanks
are not open to the atmosphere and operate above
EXPANSION
TANK
This section discusses the function and selection
of piping accessories and describes piping layout
techniques for coils, condensers, coolers, air washers,
cooling towers, pumps and expansion tanks.
ACCESSORIES
Expansion Tanks
An expansion tank is used to maintain system
ljressure by allowing the water to expand when the
water temperature increases, and by providing a
method of adding water to the system. It is normally required in a closed system but not in an
open system; the reservoir in an open system acts
as the expansion tank.
,
FIG. 29 - STRAINER LOCATION
SYSTEM
IN
W ATER PIPING
CH,\I”1’1:K 2 . W.\~I‘EK 1’Il’INO
3-31
TABLE 15-EXPANSION OF WATER
v = (0.00041 1 - 0.0‘166) v,
t
P,‘
P,
I’I
PO
(Above 40 F)
TEMP
W)
VOLUME
INCREASE
(%)
(F)
VOLUME
INCREASE
(%I
100
125
150
175
200
225
250
.6
1.2
I .a
2.8
3.5
4.5
5.6
275
300
325
350
375
400
6.0
8.3
9.8
11.5
13.0
15.0
TEMP
atmospheric pressure. Air vents must be installed
in the system to vent the air. Closed expansion tanks
arc located on the pump suction side of the system
to permit the pump suction to operate at or near
constant pressure. Locating the expansion tank at
the p~mip discharge is usually not satisEactory.
iill
‘ssure changes caused by pump operation are subtractcd from the original static pressure. If the
pressure drop below the original static is great
enough, the system pressure may drop to the boiling
point, causing unstable water circulation and possible pump cavitation. If the system pressure drops
blow atmospheric, air sucked in at the air vents can
collect in pockets and stop tvater circulation.
The capacity of a closed expansion tank is larger
than an open expansion tank operating under the
same conditions. ASME has standardized the calculation of the capacity of closed expansion tanks. The
capacity depends on whether the system is operating
above or below 160 F water temperature.
Water temperatures below 160 F use the following
formula to determine the tank capacity:
J, = E x vs
I-
P
P
2-2
Pf
p,,
where: V, = minimum capacity of the tank (gallons).
E = percent increase in the volume oE water
in the system (Table 15).
V, = total volume of water in the system
(gallons).
P, = pressure in the expansion tank when
water first enters, usually atmospheric
pressure (feet of water absolute).
P, = initial fill or minimum pressure at the
expansion tank (feet of water absolute).
P, = maximum operating pressure at the expansion tank (feet of dater absolute).
When the system water temperature is between
160 and 280 F, the following equation is used to
determine the expansion tank capacity:
where 1 = maximum average operating temp (F).
Strainers
The primary function of a strainer is to protect
the equipment. Normally strainers are placed in the
line at the inlet to pumps, control valves or other
types of equipment that should be protected against
damage. The strainer is selected for the design capacity of the system at the point where it is to be inserted in the line. Strainers for pump protection
should be not less than 40 mesh and be made of
bronze. On equipment other than pumps the manufacturer should be contacted to determine the degree of strainer protection necessary. For example, a
control valve needs greater protection than a pump
and, therefore, requires a finer mesh strainer.
Thermometers and Gages
Thermometers and gages are required in the
system wherever the design engineer considers it
important to know the water temperature or pres- :
sure. The following temperatures and pressures are
usually considered essential:
1. Water temperature entering and leaving the
cooler and condenser.
2. Pump suction and discharge pressure.
3. Spray water temperature and pressure entering
the air washer.
Water thermometers are usually selected for an
approximate range of 30 F to 200 F; they should be
equipped with separable wells and located where
they can be easily read.
Pressure gages are selected so that the normal
reading of the gage is near the midpoint of the
pressure scale.
Air Vents
Air venting is an important aspect in the design
of any water system. The major portion of the air
is normally vented thru the open expansion tank.
Air vents should be installed in the high points
of any water system which cannot vent back to the
open expansion tank. Systems using a closed expansion tank require vents at ail high points. Runoff
drains should be provided at each vent to carry
possible water leakage to a suitable drain line.
PIPING
LAYOUT
Each installation has its own problems regarding
location of equipment, interference with structural
members, water and drain locations, and provision
t
3-x
P./\K’I‘
3. I’IPINC;
DESIGN
c
lor’ service and replacement. ‘l‘he following guides
are presented to familiarize the cnginccr with accepted piping practice:
I. Shut-ofE valves are installed in the entering
and leaving piping to equipment. These are
normally gate valves. This arrangement permits servicing or replacing the equipment
without draining the entire system. Occasionally a globe valve is instaJlec1 in the system to
bcrvc as one of the shut-off valves and in addition is used to balance water flow. Most often
it is located at the primp discharge. In a close
coupled system the shut-of valves may be
omitted if the time and expense required to
drain tile system is not excessive. This is a
matter of economics, the first cost of the valves
versus the cost of new water treatment and
time spent in draining the system.
~i$Y 2. Systems using screwed, weided or soldered
joints require unions to permit removal of the
equipment for servicing or replacement. If gate
valves are used to isolate the equipment in
the system, unions are placed between the
equipment and each gate valve. Unions arc
also located before and after control valves,
and in the branch of a three-way control valve.
It is good practice to locate the control valve
between the equipment and the gate valve
used to shut,off flow to the equipment. This
permits removal of the control valve from the
system without draining the system. By locating the control valve properly, it is possible
to eliminate the unions required for removal
of the equipment. If the system &es flanged
valves and fittings, the need for unions is eliminated.
3. Strainers, thermometers and gages are normally
located between the equipment and the gate
valves used to shut off the water flow to the
equipment.
Thd following piping diagrams are illustrated
with screwed connections, However, flanged, welded
or soldered connections may be used. These layouts
have been simplified to show various principles
involved in piping practice.
Ilowing thru the coil or thru the bypass. It is rcgulatcd by a temperature controller. Gage cocks are
usually installed in both the supply and return lines
to the coil, This permits pressure gages to be connected to determine pressure drop thru the coil.
The plug cock is manually adjusted t o s e t the
pressure drop thru the coil.
Figure 31 illustrates an alternate method of piping a water coil. The plug cock shown is used to
adjust manually the water Ilow for a set pressure
drop thru the coil. The pressure drop is determined
by connecting pressure gages to the gage cocks. In
this piping layout control of the leaving air tcmperaturc from the coil is maintained within a required
range since normally the entering water is controlled
to a set temperature. Often an air bypass around the
coil is used to maintain final air temperature.
Figz~l-e 3-3 illustrates a multiple coil arrangement.
Piping connections for drain and vent lines for the
coil arc included and should be I,$ in. nominal pipe
size. The same principles covered in Figs. 30 and 31
are applicable to multiple coil arrangements.
A globe valve may be substituted for the plug
cock and gate valve combination in the return lines
in Fig. 30,31 and 32. In this arrangement the globe
valve is used to balance the pressure drop thru the
coil, and also to shut off the water when szrvicing
is required. However, it has disadvantages that
AUTOMATIC AIR VENT
THREE
pt
-WAY
Water Coils
F~~ZIWS
30 thou 36 illustrate typical piping layouts for chilled water coils in a closed piping system.
The coil layout illustrated in
30 contains a
three-way mixing valve. This valve, located at the
cooling coil outlet, maintains a desired temperature
by proportioning n?~tomnticnZZy the amount of water
NOTE: Flange or union is located
removed.’
FIG.
30--
SO
coil may be
CHILLED W ATER P IPING FOR COILS
(A U T O M A T I C: CO N T R O L )
‘
3-34
PAR.1’
3. I’II’ING
DESIGN
\
I’ig~es 34, 35 old 36 show typical piping layouts
for multiple units in a horizontal installation. The
principle difference in the three systems is the numbcr of shut-off valves (gate) and take-offs from the
hcadcr. Since the header is located under the floor,
each take-ofE must pass thru the floor. ThereEore, it
is a matter of economics to determine the number ol
shut-ofE valves required for servicing. Fig. 35 shows
the minimum number of valves that may be used,
anti Pig. 36 shows valves at each unit.
Cooler
A typical chilled water piping diagram lor a
water cooler is illustrated in Fig. 37.
In a close coupled system most of the gate valves
can be omitted. If they are omitted, all 0T the water
is drained Irom the system thru the drain valve
when a component requires servicing. In an extensive piping system the gate valves arc used to
isolate the equipment requiring servicing or replacement.
I“igu~e 37 illustrates the recommended water piping and accessories associated with a cooler.
HUT-OFF
LAYOUT
FOR
Figure 38 sliows a water-cooled condenser using
city, well or river water. The return is run higher
than tlic condenser so that the condcnscr is always
l’ull OC water. Water flow thru the condenser is modulated by the control valve in the su~~ply line.
I;igure 39 is an illustration of an alternate drain
arrangcmcnt for a con denscr discharging waste
water. Drain connections oE all types must bc
checked for compliance with local codes. Codes
usually require that a check valve be installed in
the supply lint when city water is used.
I;igzwe -/O illustrates a condenser pipecl up with a
cooling tower. If the cooling tower and condenser
arc close coupled, most of the gate valves can be
eliminated. If the piping system is extensive, the
gate valves as shown are recommended for isolating ’
the equipment when servicing.
When more than one condenser is to be used in
the same circuit, the flow thru the condensers must
be equalized as closely as possible. This is complicated by the following:
SHUT-OFF
VALVES
( GATE VALVE )
VALVE
NOTES:
1. Though not shown, control valves (automatic or
manual) may be required to control flow thru eacll
unit.
2. A shut-off valve may be installed in the supply ant1
return branch heatlers when headers serve 3 to 5
units.
3. Supply ant1 return runouts to the coil shoultl have
flared connections if the runouts are soft copper.
Otherwise unions or Hanges are installect to facilitate
servicing units.
FIG. 34 - PIPING
Condenser
HOKIZONTAL
MULTIPLE
COILS (4 UNITS - 4 S HUT-OFF V ALVES)
.
SOTES:
1. Though not shown, control valves (automatic or
manual) may he requirecl
to control flow thru each
unit.
2. Supply ant1 return runouts to the coil shoultl have
flared connections if the runouts are soft copper.
Otherwise unions or flanges are installed to facilitate
servicing units.
FIG.
35
- PINING
LAYOUT
FOR
HOKIZONTAL
MULTIPLE
COILS (4 UNITS - 2 S HUT-OFF V ALVES)
SHUT-OFF VALVE
t GATE VALVE1
NOTES:
1. Though not shown, control valves (automatic or
manual) may be required to control Row thru each
unit.
2. A shut-off valve may be installed in the supply and
return branch headers when headers serve 3 to .i
units.
3. Supply and return runouts to the coil should have
flared connections if the runouts are soft copper.
Otherwise unions or flanges. are installed to facilitate
servicing units.
FIG. 36 - PIPING LAYOUT FOR HORIZONTAL M ULTIPLE
COILS (3 UNITS - 6 SHUT-O FF V ALVES)
SUPPLY
RETURN
NOTES:
1. Flange or union is located to allow condenser head
removal.
2. With outlet at top, condenser will be flooded even
though automatic control valve is in modulating
position.
3. Check valve is required by most sanitary codes (city
water).
4. Required for city water only.
FIG. 38 - CONDENSER PIPING
S YSTEM
FOR A
O NCE -THRU
1. The pressure drops thru the condensers are
not always equal.
2. Water entering the branch line and leaving
the run of tees seldom divides equally.
3. Workmanship in the installation can affect the
pressure drop.
To equalize the water flow thru each condenser,
the pipe should be sized as follows:
1. Size the branches for a water flow of 6 fps minimum. The branch connections to each condenser should be identical.
CIRCULATING PUMP
REMOVE WELL
AND BUSHING
NOTES:
1. Flange or union is located to allow cooler head removal.
2. Gate valves shown may be eliminated in a close
coupled system.
Fro. 37 - PINING
AT A
W ATER COOLER
FOG. 39 - ;\r:rrxNArE
D RAIN CONNECTION
3-06
f'AI<'r 3. l'll'lK(; I)I:SI(;N
WATER OUT (5-10 tPS1
THERMOMETER
FOR BLOW OUT
WELL AND GUSHING
GLOW O U T T U B E S .
TO
REDUCING TEE’
NOTES:
1. Flange or union is located to allow condenser head
removal.
2. Gate valves shown may ix eliminated in a close
coupled system (except drain valve).
3. When water enters bottom of condenser, air will vent
naturally thrn cooling tower sprays. If it is necessary
to drop piping after leaving condenser, install air
vent at high point of line Ijefore drop. See dotted
line in figure.
FIG. 40 - CONI)WSER
PII~INC FOK
A
C OOLING
TOWER
HEADER-TOTAL
GPM AT 3 FPS MAX.
NOTES:
1. Thermometer wells are inserted in tees. Remove wells
to blow omit coils.
2. A single water rcgnlating valve mnst be used as
shown. If nntier capacity, install two valves in parallel
and connect pressure tribe in liquid header.
3. \Vater supply in or return ant can be at any point in
the headers.
FIG. 41 - iLIL’LTIllLE
PIPING
2. Size the
tended approsimntely
12 in. beyond the last
branch LO the condenser.
Size
the
COOLING
\vhTER
CONNECTIONS P ARALLEL)
h e a d e r lor t h e t o t a l requird water
ciuantity lor all the condensers with a velocity
OC n o t m o r e t h a n 3 E p s . T h e h e a d e r i s cs-
3.
CONDENSEK
(KEFRIGERANT
water
main
supplying
the
header lor
a velocity of 5 to 10 f p s w i t h 7 Ips a good
average. The water main may enter the header
:tt the end or at any point along the length
ol the header. Care should be used so that
crosses do not result.
4. Size the return branches, header ant1 main
in the same manner as the supply.
5. Install 2 single water regulating valve in
the main, rather than separate valves in the
hnches
to the condenser (Fig. fl).
Cooling Tower
FifiUr.efO illustrates a cooling tower ant1 condenser
piped together. Since the cooling tower is an open
piece ot’ equipment, this is 211 open piping system.
II the conclcnscr nntl cooling tower are on the same
level, a small suction head I’or the puml) exists. The
strainer should bc installed on the discharge side oE
the pump t o keep the suction side ol’ the pump ns
close to atmospheric :IS possible.
.
It is often desirable to maintain a constant water
temperature to the condenser. This is done by installing a bypass around the cooling tower. When
the condenser is at the same level or above the coolIng tower, a three-way diverting valve is recommended in the bypass section (Fig. f 2).
,A three-way mixing valve is not recommended
since it is on the pump suction side, and tends to
create VRCUUIJI conditions rather than maintain atmospheric pressures.
I;if:~e f3 illustrates the bypass layout when the
condenser is below the level of the cooling tower.
This particular piping diagram uses a two-way
automatic control valve in the bypass line. The
lriction
drop thru the bypass is sized for the unbalanced static head in the cooling tower with maxiinuniwater flow thru the bypass.
II multiple cooling towers arc to be connected,
it is recommencletl that piping be clesigned such that
the loss from the tower to the pump suction is
approximately equal for each tower. F i g . 44 illustrates typical layouts I’or multiple cooling towers.
Equ;llizing lines are usetl to maintain the same
water level in each tower.
Cl-I;\I”i’I<li
3-37
2. w.\~l‘I*:I< I’Il’Ir\‘(i
COOLING TOWER HEADERS
tir
COOLING TOWER
4y/
HEADxr=-+T~
AIf-
I
/THREE -WAY
DIVERTING VALVE
COOLING
UNBALANCED
HEAD
COOLING
/- TOWER
4,
F R O M CONDENSEI
RETURN
-BYPASS
1
FROM
,
CONDENSER
TO PUMP
SUCTION
t
____j____i
NOTES:
kN O T E 2
.4 three-way diverting valve is used when the condenser is at the same level as or above the cooling
tower. See Fig. f3 for piping layout when condenser
is below the cooling tower.
2. :\ three-way mixing valve is not recommended at this
point as it imposes additional head at the pump
suction.
1. A two-way automatic control valve is used when the
condenser is below the cooling tower. See Fig. 42 for
piping layout when the condenser is at the same
level as or above the cooling tower.
2. The friction loss from “A” to “B” includes the loss
thru that section of the pipe and the loss thru the
two-way automatic control valve. This friction loss
should be designed for the unbalanced head of the
cooling tower.
3. Locate the automatic control valve close to the cooling tower to prevent pump motor overload and tower
sptll-over when valve is. in full open position.
FIG. 42 - COOLING 71‘~w~~ PIPING FOR CONSTANT
LEAV:NG W ATER TEMI’ERATURE
(CONDENSER AND TOWER AT SAME LEVEL)
FIG. 43 - COOLING TOWER PIPING FOR CONSTANT
LEAVING W ATER TEMPERATURE
(CONDENSER BELOW TOWER)
1NO’IT.S:
I.
NOT
RECOMMENDED
TWO
COOLING
RECOMMENDED
TOWERS
RECOMMENDED
NOT RECOMMENDED
THREE
I’K;. 4-i - h’[LJLTIl’LE
COOLING TOWERS
COOLING 7‘OWEK I’II’ING
l’;\l<‘l‘ 3. I’II’ING DESIGN
338
THERMOMETER \.
GAGE -
.- ._ _
i FLOODING HEADER
SUPPLY
FLOAT VALVE
SITE ORAIN
NOTE: See
46 lrnd 47 for typical piping when an air washer is used for the dehumidifying system (section ‘LA - A”).
FIG. 45 -AIR WASHEK.PIHNC
Air Washer
The water piping layout for an air washer used
Sor humidifying is presented in Fig. 45. When the
pump and air washer are on the same level, there is
usually a small suction head available for the pump.
Therefore, if a strainer is required in the line, it
should be located on the discharge side of the pump.
Normally air washers have a permanent type screen
at the suction connection to the washer to remove
large size foreign matter.
The drain line is connected to an open-site drain
similar to those illustrated in Fig. 38 and 39. The
drain arrangement should always be checked for
compliance to local codes.
The piping layout shows a shell and tube heater
for the spray water. Occasionally heat is added by a
steam ejector instead of by a normal heater.
Chilled water is required if
accomplish dehumidification,
illustrate two typical methods
chilled water supply. The plug
the sprays are to
Figures 46 rind 47
of connecting the
cock in both dia-
DRAIN
TO
GRAVITY
GRAVITY
RETURN
TO SURGE TANK
THREE-WAY
DIVERTING
CHILLED
VALVE
WATER
RECIRCULATING
PUMP !
c’ _
l/-i
SECTION
“.A”-“A”
NOTE: Adjust plug cock so that full flow thru automatic control valve is approximately 90% of
recirculating water design.
FIG. 46 - .&R WASH!& PIPSNC USING
CONTROL VALVE
A
.I‘HKE:E-WAY
CI-I~\I”I‘ER
2 . W.\‘I‘El<
3-39
I’Il’IN(;
grams is adjusted
way diverting valve
control valve (Fig.
recirculating water
so that full llow thru the three-
(Fig. 46) and thru the automatic
47) is approximately 90% of the
design
quantity.
~;igures -fh’ and 49 are schematic sketches of multiplr air washers with gravity returns piped to the
smic header.
Sprayed Coil
SECTION
h typical layout for a sprayed coil piping system
is shown in Fig. 50. The diagram shows a water
heater which may be required for humidification.
If a preheat coil is used, the water heater may be
‘r,
eliminated.
The drain line should be fitted with a gate valve
rather than a globe valve since it is less likely to
“A’‘-“A”
NOTE: Adjust plug cock so that full Row thru diverting valve is approximately 90% of recirculating
water design.
;.
Lk5 - ,‘\IR WASILK I'II'ING USING
CONTROL
VALVE
A
become
clogged
with
sediment.
Pump Piping
TWO-WAY
?‘he following items illustrated in I’&. 5f should
be kept in mind when designing piping for a pump:
1. Keep the suction pipe short and direct.
SLOPE GRAVITY RETURN
LINES TOWARD RISER
A I R
2. Increase the suction pipe size to at least one
size larger than the pump inlet connection.
’
3. Keep the suction pipe free from air pockets.
4. Use an eccentric type reducer at the pump suction nozzle to prevent air pockets in the
V E N T ,
0
RISER
suction
UNIT
‘I
line.
4
x’:
ENTER RISER AT
SHARP ANGLE IN
DIRECTION OF FLOW
p:
4
1
\- L
IG.
48
STRAINER
ELEiATION
--AIR WASHER RETURN CONNECTIONS
DIFFERENTELEVATIONS
AT
/-
GATE VALVE
SPRAY
WATER
HEATER
-
t.
<,
-
HEADER
ENTER HEAD& AT SHARP
ANGLE IN DIRECTION OF FLOW
‘1
- SLOPE HEADER
TOWARD RISER
GATE
VALVE
NO’IX: If no strainer is installed in this location, then
a strainer is recommended on the pump discharge.
FIG-N-AIR
WASHER RETURN CONNECTIONS
!k\MELEVEL
AT
THE
FIG. Ed--SPRAYWATERCOIL
WITH
WATER
HEATER
,
I’.\l<‘L‘
3-40
3. I’II’IN<; DESIGN
x_
GAGE
/
GATE
TV A L V E @
I
I
USING TWO GAGES
GATE
.
USING ONE GAGE
-GAGE
-
GATE
FK.5i
VALVE
OR
PET COCK
-PUMP SUCTION CON'NECTIONS
,I ”
a
5. Never install a horizontal elbow at the pump
inlet. hy horizontal elbow in the suction line
should be at a lower elevation than the pump
inlet nozzle. Where possible, a vertical elbow
should lead into a pipe reducer at the pump
inlet.
If multiple pumpsare to be interconnected to the
same hcadcr, piping connections are made as illustrated in Fig. 52. This method allows each pump
FIG.
53
-GAGELOCATION
.
ATA PUMP
to handle the same water quantity. Under partial
load conditions and at reduced water flow or when
one pump is out of the line, the pumps still handle
equal water quantities.
Figz~7e 53 illustrates two methods of locating
pressure gages at the pump; one method uses two
gages and the other uses one. The use of one gage
has the advantage of always giving the correct pressure differential across the pump. Two gages may
give an incorrect pressure differential if one or both
are reading high or low.
A pulsating damper located before the pressure
gage is shown in Fig. 53. This is an inexpensive device for dampening pressure pulsations. The same
result can be obtained by using a pigtail in the line
as shown in the diagram.
Expansion
Tank
Piping
Figure 5f is a suggested piping layout for an open
expansion tank. Piping is enlarged at the connection to the expansion tank. This permits air entrained or carried along with the water to separate
I
QUICK F I L L L I N E
,-AIR V E N T
d
SIGHT
GLASS
t.
fT
G AAT ET E b
%
VALVE ’
-To- i
+I
I-’
TRAP
ENLARGED PORTION OF
RETURN LINE TO PERMIT
A I R S E P A R A T I O N ---,
‘ N O T E 2)
RETURN
LINE
,C
AT LEAST 4d
x.Ly
EXPANSION LINE
( I f ” MIN )
FIG .
E N L A R G E D T E E FOFi
AIR SEPARATION
,-NORMAL
LINE
SIZE
CIRCULATING PUMP
NOTES:
1. Do not put any valve strainer or trap in the expansion line.
2. Enlarged portion of return line and enlarged tee are
each two standard pipe sizes larger than return line.
FIG. 54 - OPEN EXPANSION TANK PII’INC
and be vented thru the tank. The expansion tank
should be located at the pump suction side at the
highest point in the system.
Vaives, strainers and traps must be omitted from
‘~‘5 expansion line since these may be accidentally
.led off or become plugged.
,-.
I;ig~re 55 illustrates the piping diagram l‘or a
closed tank.
Drain
Line
55
-
CLOSED
E XPANSION TANK
PII~INC
sure as in a blow-thru fan-coil unit. When the
system is under negative pressure as in a draw-thru
unit, the trap prevents water from hanging up in
the drain pan.
Figure 56 illustrates the trapping of a drain line
from the drain pan. The length of the water seal
or trap depends on the magnitude of the positive
or negative pressure on the drain water. For instance, a Z-inch negative fan pressure requires a
Z-inch water seal.
Normally, under-the-window fan-coil units have
the drip lyan subject to atmospheric conditions only
and the drain line from these units is not trapped.
The drain line runout for all systems is pitched
to offset. the line friction. For a single unit the runout is piped to an open site drain. Local codes and
regulations must be checked to determine proper
piping practice for an open site drain. The runout
is run full size corresponding to the drain pan
connection size.
Some applications have multiple units with the
drain lines connected to a common header or riser.
Piping
Moisture that forms on the cooling coils must be
collected and carried off as waste. On factory fabricated fan-coil units a drain pan is used to collect
this moisture. For built-up systems the floor or base
of the system (before and after the cooling coil) is
used to gather the moisture.
Since, under operating conditions, the drain water
is subject to pressure conditions slightly above or
slightly below atmospheric pressure, the line used
to carry off this water must be trapped. This trap
prevents conditioned air from entering the drain
lint when the drain water is llnder nositivr nrc+
I-
.----
-
I----
P I T C H RUNOUT
TO
OFFSET LINE FRICTION
m
m
UNIT
\
TRAP FOR
WATER SEAL
FIG. 56 - P IPING
FOR
DRAIN P ANS
342
To size the header or riser, the amount of moisture
that is expected to form niust be determined. This
moisture and the available head is used to determine
the pipe size from the friction chart for open piping systems. However, in no instance is the header
or riser sized smaller than the drain pan connection size. Also, as required in all water How
I’AKI‘
3 . 1’II’ING
DESIGN
systems, pockets
traps in the risers and mains
must be vented to prevent water hangup.
Each system should be investigated to determine
the need for drainage fittings and cleanouts for
traps. These are necessary when considerable sediment may occur in the drain pan.
.
3-43
CHAPTER 3. REFRIGERANT PIPING
GENERAL SYSTEM DESIGN
This chapter includes that practical material
quired for the design and layout of a refrigerant
piping system at air conditioning temperature levels,
using either Refrigerant 12, 22 or 500.
APPLICATION
CONSIDERATIONS
A refrigerant piping system requires the same
general design considerations as any Huid How
system. However, there are additional factors that
critically influence system design:
1. The system must be designed for minimum
pressure drop since pressure losses decrease the
thermal capacity and increase the power requirement in a refrigeration system.
2. The fluid being piped changes in state as it
circulates.
3. Since lubricating oil is miscible with Refrigerants 12, 22 and 500, some provision must
be made to:
a. Minimize the accumulation of liquid refrigerant in the compressor crankcase.
b. Return oil to the compressor at the same
rate at which it leaves.
Piping practices which accomplish these objectives
are discussed in the following pages.
CODE
REGULATIONS
System design should conform to all codes, la%
and regulations applying at the site of an installation.
addition the Safety Code for Mechanical Refrigeration (ASA-B9.1-1958) and the Code for Refrigeration Piping (ASA-B31.5-1962) are primarily
drawn up as guides to safe practice and should also
be adhered to. These two codes, as they apply ‘to refrigeration, are almost identical, and are the basis
of most municipal and state codes.
REFRIGERANT PIPING DESIGN
DESIGN
PRINCIPLES
Objectives
Refrigerant piping systems must be designed to
accomplish the following:
1. Insure proper feed to evaporators.
2. Provide practical line sizes without excessive
pressure drop.
3. Protect compressors by a. Preventing excessive lubricating oil from
being trapped in the system.
b. Minimizing the loss of lubricating oil from
the compressor at all times.
c. Preventing liquid refrigerant from entering
the compressor during operation and shutdown.
Friction Loss and Oil Return
’
In sizing refrigerant lines it is necessary to consider the optimum size with respect to economics,
friction loss and oil return. From a cost standpoint
it is desirable to select the line size as small as possible. Care must be taken, however, to select a line
size that does not cause excessive suction and discharge line pressure drop since this may result in
loss of compressor capacity and excessive hp/ton.
Too small a line size may also cause excessive
liquid line pressure drop. This can result in flashing
of liquid refrigerant which causes faulty expansion
valve operation.
The effect of excessive suction and hot gas line
pressure drop on compressor capacity and horsepower is illustrated in Table 16.
TABLE 16-COMPRESSOR
CAPACITY VS
LINE PRESSURE DROP
42 F Evaporator Temperature
COMPRESSOR
SUCTION AND HOT GAS LINE
PRESSURE
DROP
Capacity (%I
No Line loss
2F Suction Line Loss
2F Hot Gas Line Loss
4F Suction Line Loss
4F Hot Gas Line Loss
100
95.7
90.4
92.2
96.8
1 Hp/Ton
(%I
100
103.5
103.5
106.8
106.8
Pressure drop is kept to a minimum by optimum
sizing of the lines with respect to economics, making
sure that refrigerant line velocities are sufficient to
entrain and carry oil along at all loading conditions.
For Refrigerants 12, 22 and 500, consider the requirements for oil return up vertical risers.
Pressure drop in liquid lines is not as critical as
in suction and discharge lines. However, the pressure
drop should not be so excessive as to cause gas
formation in the liquid line or insufficient liquid
pressure at the liquid feed device. A system should
normally be designed so that the pressure drop in
the liquid line is not greater than one to two degrees
’
PART
3-44
change in saturation temperature. In terms of pressure drop, this corresponds to about 1.8 to 3.8 psi
for Refrigerant 12, 2.9 to 6 psi for Refrigerant 22,
and 2.2 to 4.6 psi for Refrigerant 500.
Friction pressure drop in the liquid line includes
accessories such as solenoid valve, strainer, drier and
hand valves, as well as the actual pipe and fittings
from the receiver outlet to the refrigerant feed device at the evaporator.
Pressure drop in the suction line means a loss in
system capacity because it forces the compressor to
operate at a lower suction pressure to maintain the
desired evaporator temperature. Standard practice
is to size the suction line for a pressure drop of
approximately two degrees change in saturation
temperature. In terms of pressure loss at 40 F suction
temperature, this corresponds to about 1.8 psi for
Refrigerant 12, 2.9 psi for Refrigerant 22, and 2.2
psi for Refrigerant 500.
Where a reduction in pipe size is necessary to
provide sufficient gas velocity to entrain oil upward
in vertical risers at partial loads, a greater pressure
drop is imposed at full load. To keep the total
pressure drop within the desired limit, excessive
riser loss can be offset by properly sizing the horizontal and “down” lines.
It is important to minimize the pressure loss in
hot gas lines because these losses can increase the
required compressor horsepower and decrease the
compressor capacity. It is usual practice not to exceed a pressure drop corresponding to one to two
degrees change in saturation temperature. This is
equal to about 1.8 to 3.8 psi for Refrigerant 12, 2.9
to 6 psi for Refrigerant 22, and 2.2 to 4.6 psi for
Refrigerant 500.
where h = loss of head in feet of fluid
f = friction factor
L = length of pipe in feet
D = diameter of pipe in feet
V = velocity in fps
g = acceleration of gravity = 32.17 ft/sec/sec
The friction factor depends on the roughness of
pipe surface and the Reynolds number of the fluid.
In this case the Reynolds number and the Moody
chart are used to determine the friction factor.
PIPING
DESIGN
Use of Pipe Sizing Charts
The following procedure for sizing refrigerant
piping is recommended:
1. Measure the length (in feet) of straight pipe.
2. Add 5070 to obtain a trial total equivalent
length.
3. If other than a rated friction loss is desired,
multiply the total equivalent length by the correction factor from the table following the
appropriate pipe or tubing size chart.
4. If necessary, correct for suction and condensing
temperatures.
5. Read pipe size from Charts 7 thru 21 to determine size of fittings.
6. Find equivalent length (in feet) of fittings an{
hand valves from Chapter 1 and add to the
length of straight pipe (Step 1) to obtain the
total equivalent length.
7. Correct as in Steps 3 and 4 if necessary.
8. Check pipe size.
In some cases, particularly in liquid and suction
lines, it may be necessary to find the actual pressure
drop. To do this, use the procedure described in
.
Steps 9 thru 11:
9. Convert the friction drop (F from Step 3) to
psi, using refrigerant tables or the tables in
Part 4.
10. Find the pressure drop thru automatic valves
and accessories from manufacturers’ catalogs.
If these are given in equivalent feet, change to
psi by multiplying by the ratio:
step (9)
step (6)
REFRIGERANT PIPE SIZING
Charts 7 thru 21 are used to select the proper
steel pipe and copper tubing size for the refrigeration lines. They are based on the Darcy-Weisbach
formula:
3.
11. Add Steps 9 and 10.
In systems in which automatic valves and accessories may create a relatively high pressure drop,
the line size can be increasedto minimize their effect
on the system.
Example
I
- Use of Pipe Sizing Charts
Given:
Refrigerant 12 system
Load - 46 tons
Equivalent length of piping - 65 ft
Saturated suction - 30 F
Condensing temperature - 100 F
Type L copper tubing
Find:
Suction line size for pressure drop corresponding to 2 F.
Actual pressure drop in terms of degrees F for size selected.
CHAPTER
3.
REFRIGERANT
PIPING
Solution:
Liquid Subcooling
See Chart 7.
I.
345
Line sizes for 40 F saturated suction and 105 F condensing
temperature are shown on Chart 7. Determine the correction factor for a 30 F suction temperature of 1.19 from
table in notes following Chart 9.
Where liquid subcooling is required, it is usually
accomplished by one or both of the following arrangements:
1. A liquid suction heat interchanger (heat dissipates internally to suction gas).
2. Determine adjusted tons to be used in Chart 7 by multiplying correction factor in Step 1 by load in tons:
1.19 x 46 = 55 tons
3. Enter Chart 7 and project upward from 55 tons, to a
25/8 in. OD pipe size, then to a 31/, in. OD pipe size. At
25/s in. OD, a 2 F drop is obtained with 33 ft of pipe; at
3% in. OD a 2 F drop is obtained with 71 ft of pipe.
Select a 31/s in. OD pipe to obtain less than a 2 F drop.
4. Use the following equation to determine actual pressure
drop in terms of degrees F in the 31/8 in. OD pipe with
a 46 ton load:
Actual
pressure
drop
equivalent ft of pipe
= piping allowed for 2 F drop
LINE
The amount of liquid subcooling required may be
determined by use of a nomograph, Chart 22 or by
calculation. The following examples illustrate both
methods.
Example 2 - Liquid Subcooling from Nomograph
.
’ 2 F
‘65
=71 X2=1.8F
LIQUID
2. Liquid subcooling coils in evaporative condensers and air-cooled condensers (heat dissipates externally to atmosphere).
DESIGN
Refrigeration oil is sufficiently miscible with these
refrigerants in the liquid phase to insure adequate
mixing and oil return. Therefore low liquid velocities and traps in liquid lines’ do not pose oil return
problems.
The amount of liquid line pressure drop which
can be tolerated is dependent on the number of
degrees subcooling of the liquid. Usually this
amounts to 2 F to 5 F as the liquid leaves the condenser. Liquid lines should not be sized for more
than a 2 F drop under normal circumstances. In addition, liquid lines passing thru extremely warm
spaces should be insulated.
Given:
Refrigerant 12 system
Condensing temperature - 100 F (131.6 psia)
Liquid line pressure drop (incl. liquid lift) - 29.9 psi
Find:
Amount of liquid subcooling in degrees F required to prevent flashing of liquid refrigerant.
Solution:
Use Chad 22.
1. Determine pressure at expansion valve:
131.6 - 29.9 = 101.7 psia
2. Draw line from point A (100 F cond
(101.7 psia at expansion valve).
3. Draw line from point C (intersection of AB with line 2)
thru point D (0% flash gas) to point E (intersection of
CD with liquid subcooling line).
4. Liquid subcooling at point E = 18 F. Liquid subcooling
required to prevent liquid fiashing = 18 F.
>I_
^, ,
I
..,.
cl_
-.
‘on Drop and Static Head
4th an appreciable friction drop and/or a static
head due to elevation of the liquid metering device
above the condenser, it may be necessary to resort
to some additional means of liquid subcooling to
prevent flashing in the liquid line. Increasing the
liquid line pipe size minimizes pipe friction and
flashing due to friction drop.
In large systems where the cost is warranted, a
liquid pump may be used to overcome static head.
An arrangement shown in Fig. 57 illustrates a
method which may be used to overcome the effect
of excessive flash gas caused by a high static head
in the system. This arrangement does not prevent
the forming of flash gas, but does offset the effect it
might have on the operation of the evaporator and
valves.
temp) to point B
.i
‘,.:
;.
,_.
^_&:
“.
FLOAT-ACTUATED
VENT VALVE
FIG. 57 - METHOD OF OVERCOMING ILL EFFECTS
SYSTEM HIGHSTATICHEAD
OF
3-46
PART 3. PIPING DESIGN
CHART 7-SUCTION LINES-COPPER TUBING
IEFR
/
4o”/lo50
For Pressure Drop Corresponding to 2 F
I
300
E
-200
z
WI50
E
-I
yoo
w
-1 80
a
>
3 6 0
,” 5 0
40
TONS
OF
REFRIGERATION
CHART 8-HOT GAS LINES-COPPER TUBING
For Pressure Drop Corresponding to 2 F
y
\
F
5200
\
\
I
I
i\l
I
2
3
4
\
\,
I
\
\
IR:,?
I
/ h4 \;l;‘i
\
5
,,p,343\
\
NI
II
\\5,8
J4!&
\
\
\
1
1
\
\
\
\
E 150
ifi
\
-1
+ too
\
\
\
1L
\
\
y
\
\
\
300
Y
Y
\.
\
\I \
fi 1;
t
\
\
I.
\
‘\
I
I
= 8 0
z
3 6 0
,” 5 0
5
6
8
IO
TONS
20
30 40 50 60
OF REFRIGERATION
80
100
200
3 0 0 400 500
I
(:H;\l”I‘I’K
3.
KEI:KIC;EKI\N~~‘
347
I’Il’lN(;
CHART 9-LIQUID LINES-COPPER TUBING
REFRIG. 12
For Pressure Drop Corresponding to 1 F
400/1 05O
I-----.
h
w
18
I
I
x
I\1
I
I
III1
I
\n
\ AP
I
I
,\I
l\l
I
I
\
\Illl
\I
\I
I
IUI
\
\
\I
\
2
3
4
5
6
8
20
30 40 50 60
OF REFRIGERATION
IO
TONS
Range of Chart 9:
Saturated
Suction
Condensing
Temperatures
80
-4OFto
Temperatures
I
I
200
100
Y
\
I\
300 400 500
5OF
80 F to 120 F
Pressure drop is given in equivalent degrees because of the general acceptance of this method of sizing. The corresponding pressure drop in psi may
be determined by referring to the saturated refrigerant tables.
To use Charts 7 and 8 for conditions other than 40 F saturated suction;,
105 F condensing, multiply the refrigeration load in tons
factor below and use the product in reading the chart (S = Suction, HG = Hot Gas).
-40
CO
ND
I’,P
1
-30
1
-20
SATURATED
- 1 0
1
1
SUCTION
0
TONS
S
90
100
110
120
130
140
150
160
HG
4.90 1.47
5.17 1.34
5.45 1.24
5.80 1.17
6 . 2 0 1.09
6 . 6 8 1.02
80
7.20
7.90
)
8.70
S
3.04
4.30 1.30
4.55 1.21
4.81 1.12
3.20 1.28 2.54 1.23 2.00
3.33 1.19 2.69 1.13 2.10
3.56 1.09 2.83 1.04 2.20
3.78
1.00
3.00
.98 2.37
4.03
.95 3.21
.93 2.54
.98 5.91
.95 6.43
. 9 4/7 . 1 0
1.04
.98
.94 4.35
.91 4.74
.90 1 5.20
S
HG
2.41 1.37
MULTIPLYING
S
HG
4.09 1.43
5.08
5.50
HG
1.41
.91 3.43
.88 3.74
. 8 71 4 . 0 6
TEMPERATURE
10
1
S
1.94
.88 2.71
.85 2.97
.84i 3.22
HG
1.34
S
1.60
1.22 1.68
1.12 ' 1.76
1.03 1.86
.97 1.96
.90 2.09
.86 2.24
.82 2.41
.81 12.62
by the
(F)
30
1
20
1
40
1
50
FACTOR
HG
1.32
S
1.29
HG
1.29
S
1.09
HG
1.26
1.19
1.35
1.17
1.12
1.14
1.10
1.41
1.08
1.18
1.05
1.01
1.48
1.00
1.23
.91
1.30
.85
1.39
.96
.89
.83
.94 1.58
. a 7 1.68
.83 1.80
.79 1.94
. 7 8[ 2 . 1 0
.80 1.50
.78 1.62
. 7 6 / 1.74
S
.90
.94
.98
1.02
1.06
1.15
.79 1.22
. 7 6 1.31
. 7 3 1 1.41
HG
1.24
1.12
1.03
.95
.87
.81
.77
.73
S
.77
.76
.80
.a4
.88
.96
1.03
1.11
.71 j 1 . 2 0
HG
1.22
1.10
1.01
.92
.85
.79
.75
.71
.69
NOTES:
1. To use suction and hot gas line charts for friction drop other than 2 F or liquid line charts for friction drop other than
alent length by factor below and use product in reading chart.
Friction Drop (F)
Liquid Line
0.25
0.5
Hot Gas Line
Suction Line
0.5
1 .o
4.0
2.0
Multiplier
2.
Pipe sizer ore OD and ore for Type L copper tubing.
.75
1 F, multiply equiv-
1.0
1.25
1.5
2.0
2.5
3.0
1.5
2.0
2.5
3.0
4.0
5.0
6.0
1.3
1.0
0.8
0.7
0.5
0.4
0.3
348
L
PART
?).
PIPING
DESIGN
CHART IO-SUCTION LINES-STEEL PIPE
For Pressure Drop Corresponding to 2 F
SCHEDULE
40
\
\
10
\I
III
\ I I I
- -
I
2
3
4 5 6
8
x
Y
I
\
IO
20
30
40 5060
80 100
‘(
I
200
I\I
300 400500
TONS OF REFRIGERATION
.
CHART II-HOT GAS LINES-STEEL PIPE
1REFRIG. 12 /
For Pressure Drop Corresponding to 2 F
SCHEDULE
40
h
400
300
f
k200
E
p0
w
-I
I- 100
2 80
=
5 60
," 50
TONS OF REFRIGERATION
I
I
(‘:H,\P-I‘ER
:1.
REI;RI(;ER~\NI‘
I’IPINC,
349
CHART 12-LIQUID LINES-STEEL PIPE
For Pressure Drop Corresponding to 1 F
I-
\.
y 80
3
+
S C H E D U L E 80 B
\I
h
\
SCHEDULE 40
I i\I
I I\
I\
\
\
\
I.
I
\ I
I
I I\ I
hI
I Ihlll\
h
L
I
2
3
4
5 6
8
IO
20
TONS
Range of Chart 12:
OF
30
40 5060
80 100
200
300 400 500
REFRIGERATION
Saturated Suction Temperatures
Condensing Temperatures
-40Fto
5OF
8OFto12OF
Pressure drop is given in equivalent degrees because of the general acceptance of this method of siring. The corresponding pressure drop in psi
be determined by referring to the saturated refrigerant tables.
may
To use Charts 10 and 11 for conditions other than 40 F saturated suction, 105 F condensing, multiply the refrigeration load in tons by the
factor below and use the product in reading the chart (S = S u c t i o n , H G = H o t G a s ) .
COND
TEMP
-2
-40
1
1
-30
-20
j
SATURATED SUCTION TEMPERATURE (F)
-10
/
0
10
(
20
1
30
1
40
/
50
TONS MULTIPLYING FACTOR
S
HG 1 S
HG
1 S
HG
/ S
2.26
2.36
2.46
HG
1.35
1.20
1.12
1.05
.98
.91
.87
.85
.84
1 s
1.83
1.92
2.00
2.13
2.25
2.39
2.58
2.78
3.04
1.32
1.20
1.11
1.02
.95
.89
1 S
1.54
1.59
1.69
1.78
1.89
2.00
.83
.82
.81
2.16
2.32
2.52
HG
HG I s
1.29
1.29
1.18
1.34
1.08
1.39
1 . 0 0 1.47
.93
1.56
.86
1.65
.80
.79
.78
1.78
1.92
2.08
HG ( S
1.27 1 1 . 0 7
1 . 1 6 1.12
1.08
1.17
1 . 0 0 1.23
.92
1.29
.84
1.38
.80
.78
.76
1.46
1.59
1.72
HG 1 S
1.24
.90
1.13
.94
1.05
.98
.97
1.02
.89
1.06
.82
1.13
.79
1.20
.76
1.30
.73
1.41
HG / S
1.22
.78
1.11
.79
1.02
.81
.94
.87
.87
.93
.81
.99
.77
1.06
.73
1.13
’ .71
1.23
HG
1.20
1.09
1.00
.92
.85
.79
.75
.71
.69
NOTES:
1 . T o use s u c t i o n a n d h o t g a s l i n e c h a r t s f o r f r i c t i o n d r o p o t h e r t h a n 2 F o r l i q u i d l i n e c h a r t s f o r f r i c t i o n d r o p o t h e r t h a n 1 F , m u l t i p l y e q u i v a l e n t l e n g t h b y factor b e l o w a n d u s e p r o d u c t i n r e a d i n g c h a r t .
liquid
F r i c t i o n D r o p (F)
Line
Hot Gas Line
Suction Line
Multiplier
2. Pipe sizes are nominal and ore for steel pipe.
0.25
0.5
.75
1 .o
1.25
1.5
4.0
5.0
6.0
0.5
0.4
0.3
0.5
1.0
1.5
2.0
2.5
3.0
4.0
2.0
1.3
1.0
0.8
0.7
2.0
2.5
3.0
3-50
*
PAR?’ 3. PIPING DESIGN
i
CHART 13-SUCTION LINES-COPPER TUBING
For Pressure Drop Corresponding to 2 F
i\l ill I ! I\
500
400
! I
300
IL
- 2 0 0
$50
w”
-1
blO0
w’ 8 0
;:
>
5 6 0
: 5 0
40
3c
2c
I
7
3
8
56
4
IO
3 0 4 0 5 0 6 0
20
TONS OF REFRIGERATION
8 0 100
200
300
400
500
mufirsT ‘” “fiv GAS LINES--COPPER TUBING
Drop Corresponding to 2 F
4 150
I
‘i i Ii WiiMi
\
I 11x1
\
\
\
\
5
I I III\
6
8
IO
20
30
40
Y
\
\
\
‘,
b
50 60
T O N S O F R E F R I GERATION
8 0 100
200
300 4 0 0 5 0 0
(‘;HAI”I‘ER
3. KEI’KI<;ER,\N’I‘
PIPING
,341
CHART 15-LIQUID LINES-COPPER TUBING
For Pressure Drop Corresponding to 1 F
5 0 0
\
I
I\l
I
\
I
i
I\I
4 0 0
3 0 0
i=
!!z 2 0 0
z 150
i5
J
100
s
w
8 0
2
;
6 0
c
5 0
\
\,
\
\
\,
3 0
I
I
P
u
4 0
\
\,
2
3
4
5
6
8
\,
\
\
\
\
, ._ ~ ,
-
2 0
\
\.
\,
\
IO
2 0
3 0
4 0
5 0 6 0
8 0
100
2 0 0
”
\,
TONS OF REFIGERATION
R&ge
of Chart 15:
Saturated
Suction
Condensing
Temperatures
-4OFto 5 0 F
Temperatures
8OFto12OF
Pressure drop is given in equivalent degrees because of the general acceptance of this method of sizing. The corresponding pressure drop in psi may
be determined by referring to the saturated refrigerant tables.
To use Charts 13 and 14 for conditions other than 40 F saturated suction, 105 F condensing, multiply refrigeration load in tons by factor below and
use product in reading chart. (S = Suction-HG = Hot Gas)
SATURATED SUCTION TEMPERATURE tFi
CONDENSING
- 4 0
1
- 3 0
/
- 2 0
I
- 1 0
‘WPERATURE
/
0
TON
F)
S
80
90
100
110
120
15.4
15.8
6.1
6.5
6.9
HG
/
S
1.7 / 4.3
1.6 t 4.4
1.4
4.8
1.4
5.1
1.3
5.4
HG
/
S
1.6 / 3.4
1.5 1 3.6
1.4
3.7
1.3
4.0
1.2
4.2
HG
S
HG
1.6 j 2.7
1.5 1 2.9
1.3
3.0
1.2
3.1
1.2
3.3
/
1.5
/
j
10
MULTIPLYING
S
HG
I
S
I
20
1
30
j
40
/
50
FACTOR
HG
j 2.1
1.4 1 1.7
1.4
1.4
1 2.3 1.4 1 1.8
1.3 I
1.3 ! 2.4
1.2 ’ 1.9
1.2
1.2
2.5
1.2 I 2.0
1.1
1.1
2.6
1.1
2.1
1.0
/
S
HG
/ 1.4
1.4
1.5 1.2 /
1.5
1.1
1.6
1.1
1.7
1.0
/
S
HG
/
S
1 1.1
1.3 I 0.9
1.2
1.2 1 0.9 1.1
1.2
1.1
1.0
1.3
1.0
1.0
1.3
0.9
1.1
HG
1 S
HG
1.2 / 0.8
/ 0.8 1.1
1.0
0.8
1.0
0.9
0.9
0.9
1.2
1.0
0.9
0.9
NOTES:
1. To use suction and hot gor line charts for friction drop other than 2 F or liquid line chart for friction drop other than 1 F, multiply equivalent length by factor below and use product in reodinq chart.
Friction Drop (F)
I
Liquid Line
0.25
0.5
Hot Gas Line
Suction Line
0.5
1.0
1.5
4.0
2.0
1.3
Multiplier
2. Pipe sizes ore OD and ore for Type L copper tubing.
.75
1 .o
1
(
3 0 0 4 0 0 5 0 0
1.25
1.5
2.0
2.5
3.0
2.0
2.5
3.0
4.0
5.0
6.0
1 .o
0.8
0.7
0.5
0.4
0.3
PART 3.
3-52
PIPING DESIGN
\
CHART 16-SUCTION LINES-COPPER TUBING
-l
REFRIG. 22
For Pressure Drop Corresponding to 2 F
4o”/lo50
500
400
300
c
k 200
I
I - 150
z
W
-I
F- 100
=
8 0
:
3
60
w”
50
4 0
30
20
2
I
500
3
rF
4
5
6
8
IO
30 40 50 60
T O N S OF?fEFR IGERATION
80
100
200
300 400 500
CHART 17-HOT GAS LINES-CdPPER
TUBING
For Pressure Drop Corresponding to 2 F
400
300
c
k200
E
(3 1 5 0
e
-I
c Inn
5
-1
a
>
5
E
. -80
I
6 0
50
I I\l
\I4
i\ I
\I
I
4 0
30
20
2
3
4
5
6
8
IO
20
30
40
50 60
TONS OF REFRIGERATION
8 0 100
200
300
400 500
CHAPTER
3.
REFRIGERANT
I'II'ING
3-53
CHART 18-LIQUID LINES-COPPER TUBING
For Pressure Drop Corresponding to 1 F
500
\
400
300
\
\
\I 1 I
I\II
I
I
I
N
IlIla
h
I\
I
I
I\1
!\I
I
\i
\i
I
I
I
Ihl
I
80100
200
I
I I\1
Xl
I
Y I i
I- 100
= 80
z
5 60
2
3
4 5 6
8
IO
20
TONS
kange of Chart 18:
OF
Saturated
30
Suction
Condensing
40 5060
300 400500
REFRIGERATION
Temperatures
-4OFto
Temperatures
5OF
80 F to 120 F
Pressure drop is given in equivalent degrees because of the general acceptance of this method of siring. The corresponding pressure drop in psi may
be determined by referring to the saturated refrigerant tables.
To use Charts 16 and 17 for conditions other than 40 F saturated suction, 105 F condensing, multiply the refrigeration load in tons by the
factor below and use the product in reading the chart (S = Suction, HG = Hot Gas).
SATURATED
COND
TEMP
_.
(F)
- 4 0
1
- 3 0
1
- 2 0
SUCTION
T O N S MULTJP
s
HG
s
HG
80
90
100
4.65
4.83
5.12
1.40
1.27
1.18
3.70
3.87
4.04
1.38
1.24
1.16
110
120
130
5.42 1.08
5.75 1.01
6.20
.95
4.28 1.06
4.55 1.00
4.88
.94
TEMPERATURE
(F)
- 1 0
‘INti FACT(
S
HG
s
3.39
3.60
1.04
2.70
.98 2 . 8 5
1.02
2.20
.95 2 . 3 6
1.01
.95
1.80 1.00
1.89
.91
2.01
.82
1.45
1.53
1.63
.98
.91
.83
2.14
2.27
2.46
1.73
1.85
2.02
.78
.73
.68
.77
.73
.70
s
HG
1.54 1.29 1.27 1.28
zY+&ij’ 1 . 6 1‘O1 . 1 7 ’ 1 . 3 32o1 . 1 5 ’
1.70 1.08 1.39 1.06
3o
1.21
1.28
1.36
’
4o
.97 1 . 0 1
.90 1 . 0 7
.83 1 . 1 3
’
.95
.87
.80
HG
.81
1.23
.845o 1.10
.88
1.01
.92
.96
1 .oo
.93
.87
.81
1.07
1.16
1.27
.75
.69
.64
NOTES:
1. To use suction and hot gas line charts for friction drop other than 2 F or liquid line charts for friction drop other than 1 F, multiply equivalent length by factor below and use product in reading chart.
Friction Drop (F)
Liquid Line
0.25
Hot Gas Line
Suction Line
0.5
1 .o
4.0
2.0
Multiplier
2. Pipe sizer are OD and are for Type L copper tubing.
0.5
.75
1 .o
1.25
1.5
2.0
2.5
3.0
1.5
2.0
2.5
3.0
4.0
5.0
6.0
1.3
1 .o
0.8
0.7
0.5
0.4
0.3
PAR?’ S.
3-54
PIPING DESIGN
CHART 19-SUCTION LINES-STEEL PIPE
For Pressure Drop Corresponding to 2 F
SCHEDULE
40
500
400
300
f
*.
s .LUU
. .a
I
I \I
I
z Id”
W
-I
5 100
W OF.
0 ”
.
60
50
40
00
TONS
OF
REFRIGERATION
.
CHART 20-HOT GAS LINES-STEEL PIPE
REFRIG. 22
r
500’
4o”/1 05O
-
For Pressure Drop Corresponding to
3
I
\
\
\
\
L
I\\
L
G
- 200
I
Ilhl
I I II\
\I I IIA
\I
\
\
I I\
I\,,,,
400 300
SCHEDULE 40
v
I
I\1
\
I.,\
R
I
\
l\I
2 F
\I I l\lII\ [\
I lYll\ I\
I\I
\
\I
R
\
)
\
II
..\I4
I
I
I
\3’/3
j\3” \
I\
$150
z
W
-I
+ 100
z
w
l2f-l
40
t30
I
I I’,
t-
20 lIzI
E !i
I
l\I
4
5 6
I
III
6
\
nrr41\
2’0
TONS
.
I
“lil‘k
\’
I
IO
I
OF
30
;o 50 60
REFRIGERATION
1\1II
60 100
\a
I,
I.
200
‘i
I
CHART 21-LIQUID LINES-STEEL PIPE
REFRIG. 22
For Pressure Drop Corresponding to 1 F
I
_
S
C
20
H
2
E
D
U
L
4
5
3
E
6
8
0
8
4o”/lo50
r-7
I--- SCHEDULE 40 -1
d
I
IO
20
TONS
Range of Chart 21:
OF
30
40 5060
80 100
200
I\
300 400; 00
REFRIGERATION
-4OFto
Saturated Suction Temperatures
Condensing Temperatures
50F
8 0 F to 120 F
Pressure drop is given in equivalent degrees because of the general acceptance of this method of sizing.
be determined by referring to the saturated refrigerant tablet.
The corresponding pressure drop in psi may
To use Charts 19 and 20 for conditions other than 40 F saturated suction, 105 F condensing, multiply the refrigeration load in tons by the
factor below and use the product in reading the chart (S = S u c t i o n , H G = H o t G a s ) .
3ND
_
‘4P
. )
-40
-
I
80
90
100
110
120
130
S
4.40
4.60
4.88
5.19
5.54
5.92
HG
1.38
1.26
1.18
1.09
1.00
.95
140
150
160
6.35
6.88
7.50
.90
.86
.85
-30
1
S
HG
3.53 1.36
3.71 1.22
3.91 1.14
4.14 1.06
4.40 1 .oo
4.71
.94
5.05
.90
5.45
.86
5.95
.83
-20
S
2.80
2.94
3.10
3.26
3.44
3.68
3.96
4.27
4.65
HG
1.33
1.23
1.13
1.05
.99
.92
.87
.82
.80
1
SATURATED SUCTIOk 4 T E M P E R A T U R E (F)
10
I
20
-10
1
0
TONS MULTIPLYING FACTOR
S
HG
2.28 1.31
2.40 1.20
2.52 1.10
2.66 1.01
2.81
.94
3.00
.87
S
1.86
1.93
2.01
2.13
2.26
2.40
HG
1.29
1.15
1.07
1.00
.93
.86
S
1.54
1.60
1.68
1.77
1.87
2.00
HG
1.27
1.17
1.07
1.00
.93
.85
3.20
3.43
3.75
2.58
2.75
3.01
.80
.77
.76
2.14
2.27
2.46
.79
.75
.74
.92
.80
.78
S
1.27
1.33
1.39
1.45
1.54
1.64
1.75
1.86
2.01
30
I
HG
1.26
1.15
1.05
.98
.91
.85
.79
.74
.72
S
1.06
1.10
1.15
1.21
1.28
1.37
1.46
1.54
1.67
I
HG
1.24
1.13
1.03
.95
.88
.81
.77
.73
.71
40
I
50
S
.90
.93
.97
HG
1.22
1.12
1.02 6 ~I<
S
.81
.85
.89
HG
1.21
1.11
1 .Ol
1.00
1.05
1.10
1.18
1.28
1.38
.95.
.88
.81
.75
.71
.69
- .93
.96
1.00
1.08
1.17
1.27
.93
.85
.80
.75
.70
.68
NOTES:
1. To use suction and hot gas line charts for friction drop other than 2 F or liquid line charts for friction drop other than 1 F, multiply equivalent length by factor below and use product in reading chart.
Liquid
F r i c t i o n D r o p (F)
Line
Hot Gas Line
Suction Line
Multiplier
2. Pipe sizes ore nominal and ore for steel pipe.
0.25
0.5
.75
1 .o
1.25
1.5
2.0
2.5
3.0
0.5
1 .o
I .5
2.0
2.5
3.0
4.0
5.0
6.0
4.0
2.0
1.3
I .o
0.8
0.7
0.5
0.4
0.3
CHART 22-SUBCOOLING TO COMPENSATE FOR LIQUID LINE PRESSURE DROP
-51
-4
-4,
-3
-3
”
-2
-2
:
?
-1’
-1,
00
a
,-
a
,-
REFRIGERANT
22
.
REFRIGERANT
Example 3 - Liquid Subcooling by Calculafion
Given:
Refrigerant 12 system
Condensing temperature - 100 F
Liquid lift - 35 ft
Piping friction loss - 3 psi
Losses thru valves and accessories - i:i phi
Find:
Amount of liquid sulxooling
liquid
refrigerant.
‘I‘0t:ll
pressure loss in liquid line
2. Condensing pressure at 100 F
Pressure loss in liquid line
= 116.9 psig
= 29.9 psi
Net pressure at liquid feed valve = wpsig
3. Saturation temperature at 57 psig = 82 F
(from refrigerant property tables)
4.
required to prevent flashing of
Solution:
I. Pressure loss due to pipe friction
I’ressurc loss due to valves and accessories
Pressure loss due to 35 ft liquid lift
= 3.‘,/ I.t)#
500
Subcooling required
= condensing temp - saturation temp at 87 psig
= 100 - 82 = 18 F
L i q u i d suhcooling required to prevent liquid Hashing
= 18 F.
= 3.0 psi
= 7.4 psi
= 19.5 psi
= L”3.1,
*.\t
to
1.8
of
normal liquid temperatures the static pressure loss due
elevation at the top of a liquid lift is one psi for every
ft of Refrigerant 12, 2.0 ft of Refrigerant 22, and 2.1 ft
Refrigerant 500.
Ltl.41”I‘ER 3.
Sizing
REI:RIGER~\N’I‘ PIPING
of Condenser to Receiver Lines (Condensate Lines)
piping from it conclenser to a receiver
is ru*l out horizontally (same siLc as the condcnscr
outlet connection) to allow for drainage of the contlenser. It is then droppctl vertically a sufficient distance to allow a liquid licad in the line to overcome
line friction losses. Additional head is required for
coil condensers where the receiver is vented to the
inlet of the coil. ‘I‘his additional head is equivalent
to the pressure drop across the condenser coil. The
condensate lint is then run horizontally to the receiver.
Tc~Dle I7 shows recommended sizes of the condensate line between the bottom of the liquid leg
and the receiver.
Licluid
SUCTiON
TABLE 17-CONDENSATE
CONDENSER
line
LINE DESIGN
‘;uction lines are the most critical from a design
Idpoint. The suction line must be designed to
return oil from the evaporator to the compressor
under minimum load conditions.
Oil which leaves the compressor and readily passes
thru the liquid supply lines to the evaporators is
almost completely separated from the refrigerant
vapor. In the evaporator a distillation process occurs
and continues until an equihbrium point is reached.
The result is a mixture of oil and refrigerant rich
in liquid refrigerant. Therefore the mixture which
is separated from the refrigerant vapor can be returned to the compressor only by entrainment with
the returning gas.
Oil entrainment with the return gas in a horizontal line is readily accomplished with normal
design velocities. Therefore horizontal lines can and
should be run “dead” level.
_ ,tion
3-457
Risers
i\/lost refrigeration piping systems contain a suetion riser. Oil circulating in the system can be
returned up the riser only by entrainment with the
returning gas. Oil returning lip a riser creeps up
the inner surface of the pipe. Whether the oil moves
up the inner surface is dependent upon the mass
velocity of the gas at the wall surface. The larger
the pipe diameter, the greater the velocity required
at the center of the pipe to maintain a given velocity
at the wall surface.
Tables 18, 19 and 20 show the minimum tonnages
required to insure oil return in upward flow suction
risers and the friction drop in the risers in degrees
F per 100 ft equivalent length.
Vertical risers should, therefore, be given special
analysis and should be sized for velocities that assure
TO
LINE SIZING
RECEIVER
(Based on Type L Copper Tubing)
CONDENSATE
LINE SIZE
(OD. In.)
1%
w
‘/r
REFRIGERATION, MAX. TONS
Refrigerant Refrigerant Refrigerant
22
500
12’
1.2
2.3
6.4
I
I
1
1.4
2.5
7.7
1
1.2
2.4
6.0
“X”
MIN.*
8”
1
*This is the minimum elevation required between CI condenser coil outlet and a receiver inlet for the total load when receiver is vented to
coil outlet header (bared on IO ft of horizontal pipe, 1 valve and
2 elbows).
oil return at minimum load. A riser selected 011 this
basis may be smaller in diameter than its branch
or than the suction main proper and, thcrcforc, a
relatively higher pressure drop may occur in the
riser.
This penalty should be taken into account in finding the total suction line pressure drop. The horizontal lines should bc sized to keep the total pressure
drop within practical limits.
Because modern compressors have capacity reduction features, it is often difficult to maintain the
gas velocities required to return oil upward in vertical suction risers. When the suction riser is sized to
permit oil return at the minimum operating capacity of the system, the pressure drop in this portion of
the line may be too great when operating at full
load. If a correctly sized suction riser imposes too
great a pressure drop at full load, a double suction
riser sl~ould be med (Fig. 58).
TO
COMPRESSOR
SUCTION LINE
TO
COMPRESSOR
C
ELLS
!\ --m
$00sm.
ELLS
METHOD “A”
,U-BEND
OR
2 ELLS
u
METHOD ‘8”
FIG. 58 - DOIJBLE SUCTION RISER CONSTRUCTION
l'.\l<'I‘
3-58
3.
I'II'lN(;
I)I:SIGN
\
TABLE 18-MINIMUM TONNAGE FOR OIL ENTRAINMENT UP SUCTlON RISERS
S T E E L P I P E - S T A N D A R D W E I G H T ( S C H E D U L E 40)
T = Tons of refrigeration. F = Friction Drop, degrees F per 100 ft equivalent length at tons shown.
FRICTION DROP
MULTIPLIER
3.5
7.0
12.0
18.0
‘(X/2001 1.8 x 3 . 5
300
400
500
X
‘Solve this equation to determine the friction drop multiplier for any
ratio of full load to min. load, tons.
.
TABLE 19-MINIMUM TONNAGE FOR OIL ENTRAINMENT UP SUCTION RISERS
Refrigerant
COPPER
Pipe OD
Area, Sq In.
SuctTemp
c20
+40
1%
.484
.825
T
-40
-20
0
%
F
.45
T
1.6
1%
1.256
F
.88
T
1.5
1 .56 1.0 11.1
1 .66
.6 Il.3
1.77 .A II.5
1 .89
.2 Il.8
F
T
1.5 1.4
.9 ( 1.8
.5 1 2.2
.3 12.5
.2 1 3.0
2%
3.094
1%
1.780
F
T
22
TUBING-TYPE 1
2%
4.770
F
T
3%
6.812
F
T
3%
9.213
F
T
Area,
Sq In.
?4
1
I .533
I
.8 /
.5 1
.3 t
.2 1
t -51
2.0
._
-. _
-70
--
I
+20
f40
AR
17
.-...7A .7
.87
I .o
1%
I 1.425
FIT
I
-40.-
0
1
1 .864
Suet T e m p , T
-
1
I
FIT
FIT
4.6 1.2 8.0 1.2 12.3 1.1
18.1 1.1 125.0 1.0 143.6
.6 1 21.9 .6 1 30.6 .6 1 53.0
2.8 .8 1 5.6 .7 1 9.7 .7 1 15.1
3 . 3 .5 I 6 . 7 .4 1 1 1 . 4 .4 I 17.9 .4 I 26.2 .4 3 6 . 3
.4 6 3 . 3
3.9 .31 7.8 .3113.5
.3/20.8
.2/30.8
.2 4 2 . 7 .2 7 4 . 1
A.6 .2 1 9 . 1 .2 1 1 5 . 8
.2 1 2 4 . 3 .2 135.4
.l
4 9 . 4 .I
86.5
1%
I
1 2.036
1
FIT
1 5%
1 18.67
2.3 1.3
STEEL PIPE-STANDARD WEIGHT (SCHEDULE
IPS
1 4%
1 11.97
FIT
2
1 3.355
FIT
1 2%
1 4.788
(
3
1 7.393
FIT
1 8%
1 46.85
FIT
FIT
.3
.2
.I
99.5
116.2
135.6
1
4
1 12.73
FIT
1
5
1 20.01
FIT
FIT
FIT
I.9 1 1.8 I.6 1 2 . 7 1 . 6 1 5 . 0 1 . 5 1 7 . 9 1 . 4 1 13.5 1 . 3 1 19.4 1 . 3 1 2 6 . 7 1 . 2 1 4 7 . 2 1.21 74.3
I l. l . 1 1.
I33
97
9 1 1 6--.6
.8 1 2 3 . 9
.7 1 3 3 . 0 .7 1 5 8 . 2 .71 9 2 . 1
-.- 10
..- I33
- .. 91 A7
-.. 91
.
1 .A
.7
2 . 7 .6
4 . 0 .6
7 . 4 .5
11.5
.5
1 9 . 9 .5 28.5 .5 3 9 . 2
.4 6 8 . 7 .4 110.0
T = Tons of refrigeration. F
1.6
1.8
.4
.3
3.2
3.6
.4
.2
4.7
5.3
.3
.2
8.7
10.0
.3
.2
13.6
15.7
.3
.2
23.4
26.4
.3
.2
33.6
38.6
.3
.2
46.0
53.0
= Friction Drop, degrees F per 100 ft equivalent length ot tons shown.
FRICTION DROP
MULTIPLIER
300
400
500
X
.3 199.3
.2 2 3 4 . 0
.l 2 7 2 . 0
)
6
1 28.99
1 -93.
.5
.3
F
.91 6 8 . 5 .9)137.1
.8
.61 83.6 .5 1167.0 .5
.3
.2
.l
40)
( 3’%
1 9.89
FIT
1 6’/0
1 26.83
3.5
7.0
12.0
18.0
*(x/200)‘.8 x 3.5
*Solve this equation to determine the friction drop multiplier for
ratio of full load to min. load. tons.
any
.3
.2
80.6
93.0
.3 129.1
.2 148.0
FIT
1
1
8
50.0
F
1 . 1 1148.1
1.0
.7 1182.4
.6
. A 2 1 2 . 5 .4
.3 2 5 4 . 5
.2 292.0
.2
.l
TABLE 20-MINIMUM TONNAGE FOR OIL ENTRAINMENT UP SUCTION RISERS
Refrigerant 500
Piaa
no
I
r/,
I
l’/n
1 .52
1%
. I,
IPS
?4
Area. Sa In. 1 .533
-20
I
1.8 )
.96
I
C O P P E R TUBING-TYPE
1
I 2%
I 2%
I 3%
1%
I
3%
I
4
STEEL PIPE-STANDARD WEIGHT (SCHEDULE 40)
. I,
^.I
1
I
c
I
I
1.6 1 1.9 1.5 ( 2.6 1.4 / 5.2 1.3 i 8.1 1.3 ) 13.9 1.2 120.0 1.1 1 27
= T o n s bf r e f r i g e r a t i o n . F = F r i c t i o n D r o p , d e g r e e s F p e r 1 0 0 f t e q u i v a l e n t l e n g t h a t t o n s s h o w n .
FRICTION DROP
MULTIPLIER
300
400
500
X
3.5
7.0
12.0
18.0
*(x/200)‘.8 x 3.5
*Solve this equation to determine the friction drop multiplier for
ratio of full load to min. load, tons.
Double Suction Risers
There are applications in which single suction
risers may bc sized for oil return at minimum load
without serious penalty at design load. Where single
compressors with capacity control are wised, minimum capacity corresponds to the compressor capacity at its minimum displacement. The maximum
3 minimum displacement ratio is usually three or
.ur to one, depending on compressor six.
The compressor capacity at minimum displaccmerit shoulcl be taken at an arbitrary figure of approximately 20 I; suction and not the design suction
temperature for air conditioning applications.
I\Qere multipie compressors are interconnected
and controlled so that one or more may shut down
while another continues to operate, the ratio of
maximum to minimum tlisplacemcnt becomes much
greater. In this case a double suction riser may be
necessary for good operating economy at design load.
The sizing and operation of a double suction riscl
is described as follows:
1. In Fig. 58 the minimum load riser indicated
by ‘-1 is sized so that it returns oil at the minimum possible load.
any
2. The second riser B which is usually larger than
riser A is sized so that the parallel pressure
drop thru both risers at full loacl is satisfactory,
providing this assures oil return at full load.
3. A trap is introduced between the two risers as
shown in Fig. 58. During partial load operation when the gas velocity is not sufficient to
return oil through both risers, the trap gradually fills with oil until the second riser B is
sealed off. When this occurs, the gas travels
up riser i-I only and has enough velocity to
carry oil along with it back into the horizontal
suction main.
The fittings at the bottom of the riser must be
close coupled so that the oil holding capacity of the
trap is limited to a minimum. If this is n o t done,
the t r a p cm accumulate cnougl~ oil on partial load
operation to seriously lower the compressor crnnkcase oil level. Also, larger slug-backs of oil to the
compressor occur when the trap clears out on increased load operation. Fig. TS shows that the larger
riser 11 forms an inverted loop and enters the hori/ontal suction lint from the top. The purpose of
rhis loop is to prevent oil drainage into this riser
1\JhiCll may be “iclle” during partial load operation.
Example 4 - Determination of Riser Size
Given:
Refrigerant 12 system
Type L copper tubing
Condensing temperature - 105 F
Suction temperature - 40 F
Height of riser - 10 ft
Equivalent length - 22 ft (10 ft pipe f 2 ells)
Full load - 98.5 tons
Minimum load - 8.1 tons
Two of 16 cylinders operating at 20 F
2 compressors, 8 cylinders each.
Find:
Size of riser
Suction line pressure drop
Solution:
1. From Table 18 for 8.1 tons minimum load (20 F suction
temperature) read 21/s in. OD tubing and .5 F minimum
load pressure drop per 100 ft equivalent length of pipe.
2. Calculate minimum load pressure drop:
Pressure drop for 5.6 ton
=&- X 22 = .11 F
= 145% of 5.6 tons
x 3.5 = 1.96
Minimum load pressure drop = .11 X 1.96 = .22 F
3. Determine full load pressure drop:
Full load = g = 1750% of minimum load
Full load pressure drop = .11 X 174 = 19.2 F
This full load pressure drop of 19.2 F together with a
drop for the remainder of the suction line of approximately 1.5 F results in a 20.7 F total suction line drop.
This is obviously too large and consequently a double
suction riser must be used.
4. Determine size of smaller riser (same as in Step 1 for
21/, in. OD) and large riser for suitable pressure drop
with a total load of 98.5 tons divided between the two
risers.
riser at 40 F suction temperature. Therefore, a 21/s in.
OD plus a 41/~ in. OD pipe are capable of returning oil
at maximum load.
DISCHARGE (HOT GAS) LINE DESIGN
The hot gas line should be sized for a practical
drop. The effect of pressure drop is shown
in Table 16, page 43.
pressure
Discharge Risers
Even though a low loss is desired in the hot gas
line, the line should be sized so that refrigerant gas
velocities are able to carry along entrained oil. In
the usual application this is not a problem. However, where multiple compressors are used with
capacity control, hot gas risers must be sized to carry
oil at minimum loading.
Tables 21 and 22 show the minimum tonnages
required to insure oil return in upward flow discharge risers. Friction drop in the risers in degrees
F per 100 ft equivalent length is also included.
Double Discharge Risers
Sometimes in installations of multiple compressors having capacity control a vertical hot gas
line sized to entrain oil at minimum load has an
excessive pressure drop at maximum load. In such
a case a double gas riser may be used in the same
manner as it is used in a suction line. Fig. 59 shows
the double riser principle applied to a hot gas line.
Sizing of double hot gas risers is made in the
same manner as double suction risers described
earlier.
REFRIGERANT CHARGE
Table 23 is used to determine the piping system
refrigerant charge required. The system charge
should be equal to the sum of the charges in the
refrigerant lines, compressor, evaporator, condenser
and receiver (minimum operating charge).
Let pressure drop = .5 F.
Corrected equivalent length
= equivalent length X correction factor
(notes under Chart 9)
= 22 X 4 = 88 ft (10 ft of pipe and 2 ells)
Enter Chart 7 with a 21/8 in. OD pipe and an equivalent
length of 88 ft; the capacity is 18.2 tons.
“c” /lJl
‘A”
“B*
ALTERNATE -WHERE “6’ IS
SMALLER THAN “C”
The capacity for 41/,
in. OD pipe at an equivalent
length of 88 ft is 102 tons. Therefore, a small riser of
21/, in. OD and a large riser of 41/8 in. OD in parallel
carry a load of 120.2 tons at a pressure drop of .5 F.
However, since the load is only 98.5 tons and a .5 pressure
drop is obtained with 125 ft of pipe (21/s in. pipe, 15
tons; 41/s in. pipe, 84 tons), the actual pressure drop
(degrees F) is 88/125 X .5 = 35 F.
The minimum capacity required to return oil is 34.8
tons for a 41/, in. OD riser and 6.4 tons for a 21/S in. OD
FIG . 59 - DOUBLE H OT GAS RISER
’
TABLE 21-MINIMUM TONNAGE FOR OIL ENTRAINMENT UP HOT GAS RISERS
Refrigerant 12
COPPER
TUBING-TYPE
1
TABLE 22-MINIMUM TONNAGE FOR OIL ENTRAINMENT UP HOT GAS RISERS
Refrigerant 22
COPPER TUBING-TYPE L
I 1.1
100
.081
2.1
.ll\
3.5
1 1.25 .071
2.5
.09(
4.1 .08
T = Tons of refriaeration.
.101
F = Friction Droo. deorees
5.4
.101 10.6
,091 18.1
.081 28.4
1 6.3
.081 12.5
.071 21.6
.071 33.9
F per 100 ft ecuivolent
.
.061 49
lenath at tons shown.
MIN. LOA
X
*(x/200)‘.8
x 3.5
*Solve this equation to determine the friction drop multiplier for any
ratio of full load to min. load, tons.
TABLE 23-FLUID WEIGHT OF REFRIGERANT IN PIPING
(Lb/l0 ft of length)
PIPE SIZE*
copper
OD In.
1
Steel
Nom. In.
R12
RSOO
1
1%
.013
.02 1
.043
.073
.llO
.013
.02
.042
.072
.l 1
.016
.025
.05 1
.087
.13
%
2 1%
2%
3%
1 ?h
2
2%
3
.16
.27
.42
.60
.15
.27
.41
.59
.19
.33
Sl
.72
3%
4 1%
5%
3%
4
5
.81
1.05
1.64
2.34
4.11
.80
1.04
1.62
2.33
4.06
.98
1.27
1.98
2.84
4.96
‘55
5/s
7/s
1 ?h
1%
6 ‘Vi
8%
K
%
J/4
6
8
HOT GAS LINES
100 F SAT COND TEMP
LIQUID LINES
100 F TEMP
SUCTION LINES
40 F SAT SUCT TEMP
R22
R12
RSOO
R22
.80
1.28
2.65
4.52
6.87
.70
1.13
2.33
3.98
6.06
.72
1.15
2.40
4.09
6.22
.032
.051
.105
.18
.27
.032
.052
.l 1
.18
.28
.047
.075
.16
.27
.40
9.74
16.9
26.1
37.3
8.56
14.9
23.0
32.9
8.8 1
15.3
23.6
33.7
.39
.67
1.04
1.5
.39
.69
1.1
1.5
.57
.99
1.5
2.2
2.0
2.6
4.1
5.8
10.2
2.0
2.7
4.1
6.0
10.4
3.0
3.8
6.0
8.6
15.0
R500
R12
50.5
65.5
102.0
147.0
257.0
R22
44.3
57.6
90.0
130.0
227.0
45.6
59.3
93.0
133.0
232.0
:
To Correct for Temperatures Other Than Above, Multiply by the Following Factors:
12
500
22
LIQUID
SUCT LINE-SAT. TEMP F
REFRIGERANT -
LINE-SAT.
TEMP
HOT
F
GAS
LINE-SAT.
TEMP
F
50
30
10
- 1 0
- 3 0
40
60
80
100
120
80
90
100
110
120
1.18
1.18
1.18
.84
.84
.84
.59
.58
.58
.39
.39
.39
.2b
.26
.25
1.09
1.10
1.1 1
1.06
1.07
1.08
1.03
1.04
1.04
1 .oo
1 .oo
1 .oo
.97
.96
.98
.75
.74
.74
.87
.86
.86
1 .oo
1 .oo
1.00
1.15
1.16
1.16
1.33
1.33
1.35
*Refrigerants 12, 22 and 500 weights are
for OD sizes of Type L copper pipe.
(
;
i
.I
PAR’1 3 .
3-62
in Fig. 6Oh lor
REFRIGERANT PIPING LAYOUT
cxcs
I’II’LNG D E S I G N
whcrc the compressor i.s aOo7~e
the e71trpmtor.
EVAPORATORS
Suction Line Loops
Evaporator suction lines should be laid Out to
accoml>lish the following objectives:
1. Prevent liquid refrigerant from draining into
the compressor during shutdown.
2. Prevent oil in an active evaporator from draining into an idle evaporator.
This can be done by using piping loops in the
lines connecting the evaporator, the compressor and
the condenser, Standard arrangements of suction
line loops based on standard piping practices are
illustrated in Fig. 60.
Figure 6Oa shows the compressor located below a
ogle evaporator. An inverted loop rising to the top
i the evaporator should be made in the suction line
to prevent liquid refrigerant from draining into the
compressor during shutdown.
;
A single euaporator below the compressor is illustrated in Fig. 6Ob. The inverted loop in the suction line is unnecessary since the evaporator traps
all liquid refrigerant.
Figure 6Oc shows multiple evaporators on different floor levels with the compressor below. Each
individual suction line should be looped to the top
of the evaporator before being connected into the
suction main to prevent liquid from draining into
the compressor during shutdown.
Figure 60d illustrates multiple evaporators stacked
on the same floor level or may represent a twocircuit single coil operated from one liquid’solenoid
valve with the compressor located below the evaporator. In this arrangement it is possible to use one
op to serve the purpose.
Where coil banks on the same floor level have
separate liquid solenoid valves feeding each coil,
a separate suction’riser is required from each coil,
similar to the arrangements in Fig. 6Oc and 60e for
best oil return performance. Where separate suction
risers are not possible, use the arrangement shown
in Fig. SOf.
Figure 60g shows multiple evaporators located on
the same leuel and the compressor located below the
evaporators. Each suction line is brought upward
and looped into the top of the common suction line.
The alternate arrangement shows individual suction
lines out of each evaporator dropping down into a
common suction header which then rises in a single
loop to the top of the coils before going down to
the compressor. An alternate arrangement is shown
.
When automatic compressor l~um1~clown
control
is used, evaporators are automatically kept lree of
liquid by tho pumpdown operation and, therelorc,
evaporators located a b o v e t h e compressor can be
free-draining to the compressor without protective
loops.
The small trap shown in the suction lines immediatcly after the coil suction outlet is recommended
to prevent erratic operation 0E the thermal expansion valve. The expansion valve bulb is located in
the suction line between the coil and the trap. The
trap drains the liquid from under the expansion
valve bulb during compressor shutdown, thus preventing erratic operation of the valve when the
compressor starts up again. A trap is required only
when straight runs or risers are encountered in the
suction line leaving the coil outlet. A trap is not
required when the suction line from the coil outlet
drops to the compressor or suction header immediately after the expansion valve bulb.
Suction lines should be designed so that oil from
an active evaporator does not drain into an idle
one. Fig. 60e shows multiple evaporators on diflerent
floor levels and the compressor above the evaporators. Each suction line is brought upward and
looped into the top of the common suction line if its
size is equal to the main. Otherwise it may be
brought into the side of the main. The loop prevents oil from draining down into either coil that
may be inactive.
Figure 60f shows multiple evaporators stacked
with the compressor aboue the evuporators. Oil is
prevented from draining into the lowest evaporator
because the common suction line drops below the
outlet of the lowest evaporator before entering the
suction riser.
If evaporators must be located both above and
below a common suction line, the lines are piped
as illustrated in Figs. 60~1 and bob, with (a) piping
for the evaporator above the common suction line
and (b) piping for the evaporator below the common suction line.
Multiple Circuit Coils
All but the smallest coils are arranged with multiple circuits. The length and number of circuits
are determined by the type of application. Multiple
circuit coils are supplied with refrigerant thru a
distributor which regulates the refrigerant distribution evenly among the circuits. Direct expansion
coils can be located in any position, provided proper
.
EVAPORATOR ABOVE
COMPRESSOR
STACKED ON SAME LEVELON DlFFERENT LEVELSCO MPRESSOR
ABOVE
COMPRESSOR ABOVE
(11
(al
MIJLTIPLE E V A P O R A T O R S
c, AND/OR
LOOP
ON DIFFERENT LEVELS COMPRESSOR BELOW
(cl
MULTlPLE
STACKED ON SAME LEVELCOMPRESSOR BELOW
Cdl
E”APORATORS
FIG. 60 - STANDARD ARRANGEMENTS
b
WHEN EQUAL
NECESSARY
COMPRESSOR ABOVE
COMPRESSOR BELOW
(hl
(91
MULT,P,.E EVAPORATORS ON SAME LEVEL
OF
S UCTION LINE
refrigerant distribution and continuous oil removal
facilities are provided.
In general the suction line piping principles
shown in Fig. 60 should be employed to assure
-oper expansion valve operation, oil return and
lpressor protection.
Figures 61 and 62 show direct expansion air coil
piping arrangements in which the suction connections drain the coil headers effectively. Fig. 61 shows
individual suction outlets joining into a common
suction header below the coil level. Fig. 62 illustrates an alternate method of bringing up each
suction line and looping it into the common line.
The expansion valve equalizing lines are connected
at the top of each suction header at the opposite
end from the suction connection.
Figure 63 illustrates the use of a coil having connections at the top or in the middle of each coil
header, and piped so that this connection does not
drain the evaporator. In this case oil may become
trapped in the coil. The figure shows oil drain lines
fro’m connections supplied for this purpose. The
LOOPS
(ONE-C IRCUIT COILS S HOWN)
drain lines extend from the suction connection at
the bottom end of each coil header to the common
suction header below the coil level.
EXPANSION
1”
SUCTION LINE
TO COMPRESSOR
FIG .
L O C A T E B U L B 45O
ABOVE BOTTOM OF PIPE AND
AS CLOSE AS POSSIBLE
TO COIL OUTLET
61 - DX COIL USING SUCTION CONNECTIONS
DRAIN COIL, SUCTION HEADER BELOW COIL
TO
PART 3. PIPING DESIGN
3-64
.\\
/
ALTERNATE ARRANGEMENT
SHOULD BE USED WHEN
,“A” A N D “8” A R E E Q U A L
T O “C”
w
THERMAL
r BULB
SUCTION
LINE
TO COMPRESSOR
/
FIG. 64 - DRY EXPANSION COOLER
/
L LOCATE BULB 45* A B O V E
BOTTOM OF PIPE AS CLOSE
AS POSSIBLE TO COIL OUTLET
FIG. 62 - DX COIL USING SUCTION CONNECTIONS
D RAIN COIL, SUCTION HEADER A BOVE COIL
TO
Dry Expansion Coolers
Figures 64 and 66 show typical refrigerant piping
for a dry expansion cooler and a flooded cooler respectively.
In a dry expansion chiller the refrigerant flows
thru the tubes, and the water (or liquid) to be cooled
flows transversely over the outside of the tubes. The
water or liquid flow is guided by vertical baffles.
Multi-circuit coolers should be used in systems in
which the compressor capacity can be reduced below
50%. This is recommended since oil cannot be prop
erly returned and good thermal valve control cannot
be expected below this minimum loading per circuit.
It is also recommended that the minimum capacity of a single circuit should be not less than 50yo of
its full capacity. In addition, refrigerant solenoid
valves should be used in the liquid line to each
circuit of a multi-circuit cooler in a system in which ,
the compressor capacity can be reduced below 50’7&
A liquid suction interchanger is recommended with
these coolers.
For the larger size DX coolers a pilot-operated
refrigerant feed valve connected to a small thermostatic expansion valve (Fig. 65) may be used to advantage. The thermostatic expansion valve is a pilot
device for the larger refrigerant feed valve.
.
Flooded Coolers
In a flooded cooler the refrigerant surrounds the
tubes in’the shell, and water or liquid to be cooled
flows thru the tubes in one or more passes, depending on the baffle arrangement.
Flooded coolers require a continuous liquid bleed
line from some point below the liquid level in the
cooler shell to the suction line. This continuous
bleed of refrigerant liquid and oil assures the required return of oil to the compressor. It is usually
LlPUlD S U C T I O N
INTERCHANGER
OIL
I
RETURN
VALVE
FIG. 63 - DX COIL USING OIL RETURN D RAIN
CONNECTIONS TO D RAIN O IL
FIG. 65 - HOOKUP
EXPANSION VALVE
FOR
LARGE DX COOLERS
CHAI’TEK
3.
REFRIGERANT P[I’ING
,-SOLENOID
LIQUID
SUCTION
.3-s
VALVE
BACK PRESSURE VALVE
(WHEN REQUIRED1
OUBLE SUCTION RISER
PILOT OPERATED /
VALVE
OIL BLEEDER LINE -!
F1c.66 - FLOODEDCOOLER
drained into the suction line so that the oil can be
returned to the cooler with the suction gas. This
drain line should be equipped with a hand shut-off
valve, a solenoid valve and a sight glass. The solenoid valve should be wired into the control circuit
in such a manner that it closes when the compressor
stops.
A liquid suction interchanger, installed close to
the cooler, is required to evaporate any liquid refrigerant from the refrigerant oil mixture which is
continuously bled into the suction line.
Since flooded coolers frequently operate at light
loads, double suction risers are often necessary.
To avoid freeze-up the water supply to a flooded
cooler should never be throttled and should never
bypass the cooler.
MPRESSORS
Fro67 -LAYOUTOFSUCTION
ANDHOTGASLINES
FORMULTIPLECOMPRESSOROPERATION
This allows the branch line to return oil proportionally to each of the operating compressors.
Figure 67 shows suction and hot gas header arrangements for two compressors operating in parallel.
Discharge Piping
The branch hot gas lines from the compressors
are connected into a common header. This hot gas
header is run at a level below that of the compressor
discharge connections which, for convenience, is
often at the floor. This is equivalent to the hot gas
loop for the single compressor shown in Fig. 68.
The hot gas loop accomplishes two functions:
Suction Piping
Suction piping of parallel compressors should be
designed so that all compressors run at the same
suction pressure and so that oil is returned to the
running compressors in equal proportions. To insure maintenance of proper oil levels, compressors
of unequal sizes may be erected oh foundations at
different elevations so that the recommended crankcase operating oil level is maintained at each compressor.
All suction lines are brought into a common suction header which is run full size and level above
the compressor suction inlets. Branch suction line
take-offs to the compressors are from the side of the
header and should be the same size as the header.
No reduction is made in the branch suction lines to
the compressors until the vertical drop is reached.
GAS CONNECTION
AT TOP OF
CONDENSER
HOT GAS LINE
CONDENSER
WATER-COOLED) 2
LOOP TO
FLOOR -’
COMPRESS• R-/
\y
FIG. 68 - HOT GAS LOOP
‘,
i’.\R’I‘ 3.
3-66
PIPING DESIGN
EC’UALIZER
.
FOR DRAIN
OIL AND GAS EQUALIZERS
FIG .
69 - INTERCONNECTING PIPING
1. It prevents gas, which may condense in the hot
gas line during the off cycle, from draining
back into the heads of the compressors. This
eliminates compressor damage on start-up.
2. It prevents oil, which leaves one compressor,
from draining into the head of an idle one.
Interconnecting Piping
In addition to suction and hot gas piping of parallel compressors, oil and gas equalization lines are
required between compressors and between condensing units.
An interconnecting oil equalization line is needed
between all crankcases to maintain uniform oil
levels and adequate lubrication in all compressors.
The oil equalizer may be run level with the tappings
or, for convenient access to the compressors, it may
be run at floor level (Fig. 69). Under no condition
should it be run at a level higher than that of the
compressor tappings.
Ordinarily, proper equalization takes place only
if a gas equalizing line is installed above the com-
FOR
MULTIPLE CONDENSING UNITS
pressor crankcase oil line. This line may be run level
with the tappings, or may be raised to allow head
room for convenient access. It should be piped level
and supported as required to prevent traps from
forming.
Shut-off valves should be installed in both lines
so that any one machine may be isolated for repair
without shutting down the entire system. Both lines
should be the same size as the tappings on the
largest compressor. Neither line should be run directly from one crankcase into another without
forming a U-bend or hairpin to absorb vibration.
When multiple condensing units are interconnected as shown in Fig. 69, it is necessary to equalize
the pressure in the condensers to prevent hot gas
Erom blowing thru one of the condensers and into
the liquid line. To do this a hot gas equalizer line is
installed as shown. If the piping is looped as shown,
vibration should not be a problem. The equalizer
line between units must be the same size as the
largest hot gas line.
EVAPORATIVE
CONDENSER
A
@URGE (114”) LOCATE HERF
NOT AT TOP OF
RECEIVER
I
SAFETY RELIEF
VALVE COFINECTION
( S E E ASA-89.1)
ENTIRE DRAIN LINE
TO BE SAME SIZE,
AS COIL OUTLET
&‘HORIZONTAL
LENGTH OF CONDENSATE
PIPING LESS THAN 6 FEET 1 MUST
CONNECT TO UPPER PART OF RECEIVER1
FIG. 70 - HOT GAS A ND LIQUID PIPING, SINGLE COIL UNIT W ITHOUT RECEIVER VENT
CONDENSERS
Liquid receivers are often used in systems having
evapl>rative or air-cooled condensers and also with
V’
r-cooled condensers where additional liquid
su,‘age capacity is required to pump down the
system. However, in many water-cooled condenser
systems the condenser serves also as a receiver if the
total refrigerant in the system does not exceed its
storage capacity.
When receivers are used, liquid piping from the
condenser to the receiver is designed to allow free
drainage of liquid from the condenser at all times.
This is possible only if the pressure in the receiver
is not allowed to rise to the point where a restriction
in flow can occur.
Liquid flow from the condenser to the receiver
can be restricted by any of the following conditions:
1. Gas binding of the receiver.
2. Excessive friction drop in the condensate line.
3. Incorrect condensate line design.
The following piping recommendations are made
to overcome these difficulties.
Evaporative Condenser to Receiver Piping
Liquid receivers are used on evaporative condensers to accommodate fluctuations in refrigerant
liquid level, to maintain a seal, and to provide storage facilities for pumpdown. An equalizing line
from the receiver to the condenser is required to
relieve gas pressure tending to develop in the receiver. Otherwise liquid hang-up in the condenser
due to restricted drainage can occur. The receiver
can be vented directly thru the condensate line to
the condenser outlet, or by an external equalizer
line to the condenser.
Figure 70 shows a single evaporative condenser
and receiver vented back thru the condensate drain
line to the condensing coil outlet. Such an arrangement is applicable to a close coupled system. A
separate vent is not required. However, it is limited
to a horizontal length of condensate line of less than
3-68
L
PART 3. PIPING DESIGN
VENT LINETO
OUTLET
HEADER
/ COIL
r
/
EVAPORATIVE
I
CONDENSER
\
DETAIL
HOT GAS LINE -
TO
\
“Y”
EVAPORATOR
/
I
f
.
AT TOP OF RECEIVER
RECEIVER VENTfSIZE FROM TABLE 241
TO COIL OUTLET HEADER
4
I /’
SAME ,I!, AS
TO SECOND ELBOW
FROM TABLE 17
FOR VENTING AS
SIZE CONDENSATE LINE
SHOWN.OTHERWISE
(FROM TABLE I7 1
CONSULT
CONDENSER MANUFACTURER
FIG . 71 - HOT
GAS
AND
SAFETY RELIEF
I - - - - VALVE CONNECTION
&
\(SEE ASA-B9.11
,,&j
’ ‘B”~‘;O;~~‘E”E”~‘::;‘:;OVE
L IF AT BOTTOM;‘X”MIN. WOULD
BE MEASURED FROM COIL
OUTLET TO BOTTOM
CCNNECTION 1
LIQUID P IPING, S INGLE COIL UNIT WITH RE CEIVER V ENT
6 ft. The entire condensate line from the condenser
to the receiver is the same size as the coil outlet.
The line should be pitched as shown.
Figure 71 shows the refrigerant piping for a single
unit with receiver vent. Note that the condensate
line from the condenser is the full size of the outlet connection and is not reduced until the second
elbow is reached. This arrangement prevents trapping of liquid in the condenser coil.
Table 24 lists recommended sizing of receiver
vent lines.
There are some systems in current use without a
receiver but it must be recognized that problems can
occur which can be avoided if a receiver is used.
Such an arrangement is more critical with respect
to refrigerant charge. An overcharged system can
,
waste power and cause a loss of capacity if the overcharge backs up into the condenser. An undercharged system also wastes power and causes a loss
of capacity because the evaporator is being fed partially with hot gas. Therefore, if the receiver is
omitted, an accurate refrigerant charge must be
maintained to assure normal operation.
TABLE 24-RECEIVER VENT LINE SIZING
Receiver to Condenser
VENT LINE SIZE BASED ON
TYPE 1 COPPER TUBING
(In. OD)
REFRIGERATION
(Tons, Max.)
“‘8
.I’.F
NOTE:
MULTIPLE CONDENSER COMBINATIONS
HAVE SAME COIL CIRCUIT LENGTHS,
PURGE l/4” LOCATE HERE NOT
AT TOP OF RECEIVER
--..--..--..\
/
----A
/‘I.
RECEIVER VENT
(SIZE FROM TABLE 2.4)
TO COIL OUTLET HEADER
SAFETY RELIEF
VALVE CONNECTION
(SEE ASA-89.1)
LIOUID LEVEL SIGHT GLASS
(OPTIONAL)
FOR TOTAL LOAO ---GLASS
i
w
SAME SIZE AS OUTLET
TO FIRST TEE
FIG.
72 -
LOAO ( FROM TABLE 17)
INLET MAY BE LOCATED AT
GOTTOM SEE FIG. 71 OETAIL “Y”
I IF AT BOTTOM “X” MIN WOULD
SIDE TO AVOID
FORMING TRAP
‘VALVE NO. I IS OPTIONAL PROVlOlNG
BOTH VALVES NO. 2 ARE USED
HOTGAS AND LIQUID PIPINC,MULTIPLEDOUBLE COILUNIT
The advantage of such an arrangement is an ecoIx
c one; equipment cost is lower since receiver
anhvalves are eliminated and the system operating
charge is lower if charged accurately.
Figure 73 shows a subcooling coil piping ‘arrangement. The subcooling coil must be piped between
the receiver and the evaporator for best liquid subcooling benefits.
Figure 72 shows a piping arrangement for multiple units. Note that there are individual hot gas
and vent valves for each unit. These valves permit
operation of one unit while the other is shut down.
These are essential because otherwise the idle unit,
at lower pressure, causes hot gas to blow thru the
operating unit into the. liquid line. Purge cocks
are also shown, one for each unit.
Multiple Shell and Tube Condensers
The hot gas piping should be such that the
pressure in each condenser is substantially the same.
To accomplish this the branch connection from the
hot gas header into each condenser should be the
same size as the condenser coil connection.
When two or more shell and tube condensers are
applied in parallel in a single system, they should
be equalized on the hot gas side and arranged as
shown in Fig. SF.
The elevation difference between the outlet of
the condenser and the horizontal liquid header must
be at least 12 in., preferably greater, to prevent gas
blowing thru. The bottoms of all condensers should
be at the same level to prevent backing liquid into
the lowest condenser.
When water-cooled condensers are interconnected
ns shown, they should be fed from a common water
regulating valve, if used.
I’ART
3-70
3. PIPING D E S I G N
HOT GAS LINE
--)
LIQUID L I N E
TO EVAPORATOR
,I-
CONDENSER COIL
SUBCOOLING COIL
CONNECTIONS -
TURN VALVE ON
SIDE TO AVOID
SECOND
ELBOW
FIG . 73 - S UBCOOLING COIL P IPING
An inverted loop of at least 6 ft is recommended
in the liquid line to prevent siphoning of the liquid
into the evaporator (or evaporators) during shutdown. Where a liquid line solenoid va!ve or valves
are used, the loop is unnecessary.
Figure 74 shows a similar loop for a single condenser with the evaporator below.
6’ MINIMUM
Vibration of Piping
Vibration transmitted thru or generated in refrig
erant piping and the objectionable noise which results can be eliminated or greatly minimized by
proper design and support of the piping.
The best way to prevent compressor vibration
from being transmitted to the piping is to run the
suction and discharge lines at least 6 pipe diameters
in each of three directions before reaching the first
point of support. In this manner the piping can absorb the vibration without being overstressed.
LIQUID LINE
/
II
SUCTION FROM EVAPORATOR
FIG . 74
TO EVAPORATOR BELOW
- LIQUID LINE FROM CONDENSER OR RECEIVER
TO E VAPORATOR L OCATED B ELOW
\
TO EVAPORATOR
I
( I F ABOVE
C O N D E N S E R S I-4
CONDENSERS
OR
RECEIVERS
TO
EVAPORATOR
CONDENSERS)
LEG
FIG.
55 - L IQUID
PIPING
TO
INSURE CONDENSATE F LOW F ROM INTERCONNECTED CONDENSERS
The hot gas loop from the compressor can be
attached to the compressor base by means of a
bracket il the base is isolated. If there is enough
space in the horizontal run of the loop, two brackets
are recommended to eliminate excessive rocking
mc-‘ement of the pipe. Brackets should be attached
a b :e point of minimum movement of the compressor assembly. The riser following the loop is
supported as close as possible to the compressor.
If the compressor is mounted on a resilient base,
the pipe support should have a resilient isolator.
The isolator is selected for four times the deflection
in the spring support of the compressor base.
S e e “Vihntion Isohtion of Piping Systems” in
Cjicrptet- I for further discussion of the subject,
REFRIGERANT PIPING ACCESSORIES
LIQUID
LINE
liquid Suction lnterchbngers
These are devices which subcool
the liquid refrigerant and superheat the suction gas. The follow-
ing describes four reasons for their use and the best
location for each application:
1. To subcool the liquid refrigerant to compensate for excessive liquid line pressure drop.
Location - near condenser. Liquid suction interchangers are not recommended for single
stage applications using Refrigerant 22 because
superheating of the suction gas must be limited to avoid compressor overheating. However, where they are used to prevent liquid
slop-over to the compressor, superheating of the
suction gas should be limited to 20 F above
saturation temperature. A IiqUid suction interchanger so designed to limit the superheat of
the suction gas should have a bypass so that
operating adjustments may be made.
1. To act as an oil rectifier. Location - near cvaporator.
3. io prevent liquid slop-over to the compressor.
Location - near evaporator.
-1. To increase the effciency of the Kefrigerant
I2 and 500 cycles. Location - near evaporator
to avoid insulation 0E subcooled liquid line.
I’:\R’r 3. P[I’ING DESIGN
3-72
GAS IN
INTERCHANGER TEE
OBTAINABLE
FOR
ANY
,
CONDITION
LIQUID
L,Q”lD O U T
TEE,
NOTE :
SlZE
THE OUTER PIPE
ONE
PIPE
THE
,NNER
SIZE
IN
PREFERRED CONSTRUCTION
T BUSHING
(LIOUIO)
LARGER
THAN
PIPE (SUCTION).
.
ALTERNATE
CONSTRUCTION
FIG. 76 - LIQUID SUCTION INTERCHANGER
.
Two common types of liquid suction interchangers are:
1. The shell and coil or the shell and tube exchanger, suitable for increasing cycle efficiency
and for liquid subcooling. This type is usually
installed so that the suction outlet drains the
shell to prevent oil trapping.
2. The tube-in-tube interchnnger (Figs. 76 and
77), a preferable type for controlling slop-over
caused by erratic expansion valve feed or for
“rectifying” lube oil from a refrigerant oil
mixture bled from a flooded evaporator,
Excessive superheating of the suction gas must be
avoided with heat exchangers since it causes excessive compressor discharge temperatures. Therefore,
the amount of liquid subcooling permissible by a
liquid suction interchanger is limited to the amount
of suction gas superheating that does not cause compressor damage when the gas is compressed to the
discharge pressure. Beyond this point additional
subcooling should be obtained from other sources.
Charts 23 and 2-f are used to determine the length
(A) of a concentric tube-in-tube interchanger (Fig.
76). The amount of liquid subcooling available is
SUCTION
GAS OUT
SUCTION
GAS IN
LIQUID OUT
LIQUID IN
FIG. 77 - E CCENTRIC THREE -PIPE LIQUID SUCTION INTERCHANGER
.
CHART 23-EFFICIENCYCURVES, DOUBLE TUBE LIQUID SUCTION INTERCHANGER,
Refrigerants
12
and
500
SUCTION
TUBE SIZES
LENGTH OF INTERCHANGER (FTl”A”
CHART
24-EFFICIENCY
CURVES,
DOUBLE TUBE
Refrigerant
LIQUID
22
a3 0
2
?I
$
ii
::
,
iA
20
IO
LENGTH OF INTERCHANGER (FT) “A”
SUCTION
INTERCHANGER,
s-74
l’.ZK’I-
L
3.
PIPING
DESIGN
by using the ratio of the specific heats of
the suction gas and of the liquid (subcooling multiplier). Example 5 illustrates the use of these charts.
calculated
Example 5 - Determining the Length
Tube-in-Tube
of a Concentric
Interchanger
Given:
Refrigerant 12 system
Load - 45 tons
Suction line - Si/, in. OD Type L copper
Expansion valve - 10 F superheat
Suction temp - 40 F
Condensing temp - 105 F
Find:
Length of a concentric tube-in-tube interchanger to superheat the suction gas to 6.5 F (suction gas temperature to
compressor in accordance with ASRE Standard 23R on
rating compressors).
Amount of liquid subcooling.
Solution:
See Chart23.
1.
Determine
E=
.
FIG.
interchanger
efficiency
E
from
PoRTL~QUID
INDICATOR
equation
leav gas temp - ent gas temp
ent liq temp - ent gas temp x 100
= s x 100 = +X 100 = 27.2%
2. With an efficiency of 27.2% a 31/8 in. OD pipe size and
a 45 ton load, enter Chart Zjr as indicated per dashed
line to determine a length (A) of 17 ft.
3. For Refrigerant 12 the ratio of gas to liquid specific heat
is ,653. Therefore, subcooling of liquid refrigerant is
,653 X 15 F (leav gas temp - ent gas temp) or 9.8 F.
An eccentric three-pipe interchanger is shown in
Fig. 77. The inner pipe and the outer pipe offer two
surfaces for the exchange of heat between the warm
liquid refrigerant and the colder suction gas. The
required length of this interchanger can be determined by using the method shown in Example 5
and by basing the required length on a ratio of relative surfaces between the liquid refrigerant and the
suction gas.
liquid indicators
Every refrigeration system should include a means
of checking for sufficient refrigerant charge. The
common devices used are a liquid line sight glass,
liquid IeveI test cock on condenser or receiver, or
an external gage glass with equalizing connections
and shut-off valves.
The liquid line sight glass is one of the most
convenient to install and use. A properly installed
sight glass shows bubbling when there is an insufficient charge and a solid clear glass when there is
sufficient charge.
Sight glasses should be installed full size in the
main liquid line and not in a bypass line that parallels the main line.
.
78 -DOUBLE
A sight glass with double ports and seal caps is
preferable. The double ports allow a light to be put
behind one port so that the state of the refrigerant
is easily seen. The seal caps serve as an added protection against leakage or breakage since they are
removed only when checking the refrigerant. Fig. 78
shows a double port liquid indicator with seal caps.
Theinstallation of a double port or see-through
sight glass is recommended in each of the following
locations:
1. On evaporative condensing installations - in
the liquid line leaving the receiver.
2. On single water-cooled condenser installations
- in the liquid line leaving the condenser or, if
an auxiliary receiver is used, in the liquid line
leaving the receiver.
3. On multiple water-cooled condenser installations - in the main liquid line leaving the
bank of condensers and also in the liquid line
leaving the receiver if an auxiliary receiver is
used.
Strainers
The installation of a strainer ahead of each automatic valve is recommended. Where multiple expansion valves with integral strainers are used, a
single main liquid line strainer is sufficient to
protect all of these. Fig. 79 shows an angle cartridge
type strainer. A shut-off valve on each side of the
strainer is desirable and should be located as close
to the strainer as possible.
On steel piping systems an adequate strainer
should be installed in the suction line and a filterFigures
78-80,
courtesy of ~Mueller Brass CO.
FIG. 80 - ~INGI.IC
FIG.
59
-
TYPE DKIEK-STKAINKH
ANGLE CARTRIDGE T YPE S TRAINER
drier in the liquid line to remove the scale and rust
inherent in steel pipe.
Refrigerant Driers
A permanent refrigerant drier is recommended
‘lost systems and is essential for all low temperafr
tL,.i systems. It is also essential for all systems using
hermetic compressors since the compressor motor
winding is exposed to refrigerant gas. If the gas contains excessive moisture, the winding insulation may
break down and cause the motor to burn out. A fullflow drier must be used for this type system.
Figure 80 shows an angle type cartridge drier.
The drier should be mounted vertically in the
liquid line near the liquid receiver. A three-valve
bypass (Fig. 81) should be used to permit isolation
of the drier for servicing and to allow partial refrigerant How thru the drier.
-r
FIG. 8 1 - T HREE-V ALVE B YPASS
DRIER
FOR
REFRIGERANT
DRIER
Reliable rnoistzlre indicator-s (Fig. 82) for liquid
refrigerant lines are available. These devices indicate the proper time to replace the drier cartridge.
Filter-Driers
Filter-driers (Fig. 83) are more commonly used
t’ 1 strainers and driers together. The drier mater;U1 actually filters the liquid refrigerant.
Solenoid Valves
Solenoid valves are commonly used in the following places:
1. In the liquid line of any system operating on
single pump-out or pump-down compressor
control.
2. In the liquid line of any single or multiple
DX evaporator system.
3. In the oil bleeder lines from flooded evaporators
to stop the flow of oil and refrigerant into
the suction line when the system shuts down.
In many cases it is desirable to use solenoid valves
with opening stems. The opening stem serves as a
by-pass so that the system may continue to operate
in case of solenoid coil failure.
FIG.
82
-
COMBINATION M OISTURE
AND
L IQUID
INDICATOR
FIG.
83
- FILTER-DRIER
Figures 82 and 83. COU~WS~ of Sporlnn Vnlve C O .
I’,\K’I‘
3-TG
S. PIPING D E S I G N
Five degrees
is the usual change in superheat between a full open and closed position. This
is called the operating superheat. Thus a valve
which operates at 10 degrees superheat at design
load balances out at 5 to 6 degrees superheat at low
load. A low superheat setting at design load, therefore, does not provide sufficient margins of safety at
low loads because of the 5 degrees
necessary for
operating superheat.
The expansion valve bulb should be located immediately after the coil outlet on the suction line
and 45”above
the bottom of the pipe. With this
arrangement the coil is the source of superheat
for valve operation. The valve should be set so that
overfeeding does not occur at times of partial load.
heat.
F I G . 8 4 - I<EFKLGERANT C H A R G I N G C O N N E C T I O N S
Refrigerant Charging Connections
The two usual methods of charging the refrigeration system are:
1. Charging liquid into the liquid line between
the receiver shut-off valve and the expansion
valve. Fig. 8-f shows a charging connection in
a liquid line leaving a receiver.
2. Charging gas into the suction line. Except for
very small systems this method is not practical
because of the time required to evaporate the
refrigerant from the drum and because of the
danger of dumping raw liquid into the compressor.
Expansion Valves
Thermal expansion valves should be sized to
avoid both the penalties of being undersized at full
load and of being excessively oversized at partial
load. The following items should be considered before sizing valves:
1. Refrigerant pressure drop thru the system must
be properly evaluated to determine the correct
pressure drop available across the valve.
2. Variations in condensing pressure during operation affect valve pressure and capacity. Condensing pressure should be controlled, therefore, to maintain required valve capacity.
3 . Oversized thermal expansion valves do not
control as well at full system capacity as properly sized valves and control gets progressively
worse as the coil load decreases. Capacity reduction, available in most compressors, further
increases this problem and necessitates closer
selection of expansion valves to match realistic
loads.
When sizing thermal expansion valves, make the
selection on the basis of maximum load at the design operating pressure and at least 10 degrees super-
The effect of condensing temperature on the
capacity of an expansion valve for two differentl
systems is illustrated in Example 6.
Example 6 - Effect of Condensing Temperature on
Expansion Valves
Given:
Two refrigeration systems using Refrigerant 500, one operating at 40 F suction and 90 F condensing, the other operating at 40 F suction and 130 F condensing.
218.2 psig
6.2
Coxidensing pressure
Liquid line drop
Pressure at thermal
expansion valve inlet
Suction pressure
Suction line losses
Coil pressure drop
Distributor pressure
212.0 psig
46.2 psig
2.8
7.0
17.0
drop
Pressure at thermal
expansion valve outlet
Pressure drop available
across valve
73.0 psig
139.0 psi
Assume that a valve of 27.5 ton capacity at 40 F suction
and 60 psi differential is selected.
Find:
Capacity at the pressure drop available across the valve of
systems 1 and 2.
Solution:
The capacities will
of the pressures:
vary
approximately
as
the
square
root
1%
For system 1
Cap. T 23 tons
For system 2
Cap. = 42 tons
Note that the expansion valve capacity is nearly double at
the higher head pressure.
<;H;\YI‘EK
3.
KEFKIGEKANT
l?II’INC;
On certain low temperature applications and on
high temperature applications where the design or
partial load least temperature difference (L.T.D.)
between the refrigerant and air or water is extremely
small, it may become necessary to consider the use
of the liquid suction interchanger as a source of
superheat. This is done to increase the effective
evaporator surface by allowing the liquid suction
interchanger to supply the superheat function.
If only one liquid suction interchanger is used
for the applications just mentioned, it should be an
eccentric three-pipe interchanger as shown in Fig.
77. This arrangement permits the expansion valve
bulb to sense the suction gas temperature from the
outside surface of the interchanger. Otherwise two
tube-in-tube interchangers should be used with the
thermal expansion valve bulb located between the
iv -rchangers.
,‘he preferred refrigerant flow in a coil circuit to
obtain superheat is illustrated in Fig. 55.
SUCTION
LINE
Back Pressure Valves
A conventional type back pressure regulating
valve is used in a refrigerating system to maintain
a predetermined pressure in the evaporator. A conventional type regulator controls the upstream pressure. The regulator has a spring loaded diaphragm
designed to actuate a seat pilot valve. The actuating
pressure comes from the evaporator or upstream
side of the regulator. When the upstream pressure
against the diaphragm is greater than that exerted
by the spring, the pilot valve opens and a flow of gas
is admitted to the power piston. The piston in turn
causes the main port to open. This permits a flow of
from the upstream side of the valve to the down‘b-.;earn side. When the actuating pressure becomes
less than that controlled by the spring pressure, the
flow of gas to the power piston is stopped and the
regulator closes.
There are many variations of the back pressure
regulating valve. Several are described in the following:
1. The compensating type, actuated by air or
electricity, varies the suction pressure in accordance with temperature or load demand.
2. The dual pressure regulator is designed to
operate at two predetermined pressures without‘ resetting or adjustment; by opening and
closing a pilot solenoid, either the low pressure
or the high pressure head operates.
Figzlre 86 shows a simple back pressure regulating
3-77
-
AIR
REFRIGERANT7
FN;. 85 - ~‘RI-IlXK1’:I)
C1RC:I:I-r TO OIsl~AIN
~I.I:RI(;ICRA\NT
SuI’I<RHI:.A2.L.
FLOW
IN COIL
(l’I.AN VIEW)
valve which is ordinarily used for one of the following purposes:
1. To control evaporator suction pressure in spite
of compressor suction pressure variation.
2. To establish evaporator suction pressure when
lower compressor suction pressure is demanded
by another part of the same system.
3. To prevent evaporator freezing when operating near the freezing temperature.
PRESSURE
ADJUSTING
STEM
GAGE
CONNECTION -
/
LCOPPER TUBE
CONNECTION
FOR REFRIGERANTS
12, 22 a 500
OPENING STEM
‘SEAL CAP B PACKING GLAND
FIG. 86 - BACK
P RESSURE V ALVE
.
P,\K’I‘ 3. PIPING
3-78
CHART 25-BACK PRESSURE VALVE APPLiCATlON
CHART
EXPANSION
FIG.
87 -INSTALLATION
USING BACK P RESSURE V ALVES
DI3IGN
CHAI’TEK 3.
KEI~K1~~EK,\N’I‘
3-79
I’II’LNC
C/WJ.~ 135 illustrates the application ot back pressure valves for various services such as number and
types of evaporators, and types of room and comprcssor control.
[;igure 87 illustrates the location of back pressure
valves.
DISCHARGE LINE
Oil
Separators
Oil separators reduce the rate of oil circulation.
However they are not 100CJ$ efficient since some oil
always circulates thru the system.
Oil separators are of particular value on certain
types of installations such as:
1. Systems requiring a sudden and frequent
capacity variation.
2. Systems having extensive pipe lines and numerous evaporators. The large volumes inherent
in s&h systems result in appreciable oil
hangup.
‘I-here are several objections to oil separators:
1. Oil separators permit some oil to be carried
over into the system and, therefore, proper piping design to return oil is still required even
though a separator is used.
2. On start-up, gas may condense in the shell of
the separator. As a result the separator delivers
liquid refrigerant into the crankcase. This in
turn increases crankcase foaming and oil loss
from the compressor.
During the “off” cycle the oil separator cools
down and acts as a condenser for liquid refrigerant
that evaporates in the Garmer parts of the system.
Thus a cool oil separator acts as a liquid condenser
during “off” cycles and also on compressor start-up
until the separator has warmed up. Large amounts
i ,quid refrigerant in the crankcase result in poor
lubrication and may also result in removing the oil
from the crankcase completely.
Figwe 88 shows the recommended method for
piping an oil separator.
F1c;.88 - OIL SEPARATOR
LOCATION
Check Valves
Check valves contribute a relatively large addition to a line pressure drop at full load and must
be taken into account in the selection of refrigeration cquipmcnt. In addition a check valve cannot be
relied upon for 100% shut-off.
Whenever the receiver is warmer than the compressor during shutdown, refrigerant in the receiver
tends to boil off and flow back thru the condenser
and hot gas discharge line to the compressor where
it condenses. If thcrc is sufficient refrigerant in the
receiver, liquid refrigerant eventually enters the
compressor despite the loop in the hot gas line at
the base of the compressor. To prevent this, a check
valve should be used (Fig. 68, page 65).
In a non-automatic system a hand valve may be
used at the inlet to the condenser to manually shut
off the flow on shutdown, in which case the pressure
drop involved will be much less than that cncountcred using a check valve.
REFRIGERANT
PIPING
INSULATION
Liquid lives should not be insulated if the surrounding temperature is lower than or equal to
the temperature of the liquid. Insulation is recom-
, HOT GAS LINE
Mufflers
II a muffler is used in the hot gas line, it should
be installed in downward flow risers or in horizontal
lines as close to the compressor as possible.
The hot gas pulsations from the compressor can
set up a condition of resonance with certain lengths
of refrigerant piping in the hot gas line. A muffler
installed in the compressor discharge aids in climinating such a condition.
FigWe 89 shows a muffler in a hot gas line at
the compressor.
\
WEDGE CAP
BOTH ENDS
HOT GAS MUFFLER
COMPRESSOR
-!
3-N
n~entlccl only whl the liciuid line can pick L I P a
consitleral~lc
amount ol heat. The lollowing areas
in a refrigerant piping system should be insulated:
I. .\ liquid line exposed to the direct rays ol the
S L I P lor ;I considerable distance.
2. Piping in boiler rooms.
3. Piping at the outlet of a liquid suction interchanger to preserve the subcooling effect.
Where liquid and suction lines cm be strapped
togcthcr, a single insulating covering cm be used
over both lines. This inducts an exchange of heat
ant1 is desirable from the standpoint of the subcooling effect on the liquid. However excessive superheating of the suction gas can result from too much
exchange oE heat.
Hot gns li’nes should not be insulated. Any heat
lost by these lines reduces the work to be done by the
condenser.
Sztclion lines should be insulated only to prevent
dripping where this causes a nuisance or damage.
It is generally desirable to have the suction line
capable of absorbing some heat to evaporate any
liquid which may have entered the suction line
from the evaporator. For unusual conditions of
Iligh ambient temperatures and simultaneous high
PAR’1
3.
I’LI’ING
DESLGN
relative humidities extra insulation must be applied.
The thickness oE insulation required to prevent
condensation on the outer surlacc is that thickness
which raises the temperature of the outer surlacc
ol’ the insulation to a point slightly higher than
the m;iximum expected dewpoint of the surrounding air. The external vapor barrier must be made
as nearly perfect as possible in order to prevent
leakage of vapor into the insulation.
Kcgular “ice water” thickness moulded cork covering wired on and sealed with asphalt primer is
desirable l’or most work in the air conditioning
range. For lower temperatures, “brine” thickness
moulded cork covering should be used. ‘Insulation
that is not vapor-proof soon becomes saturated with
moisture and rapidly deteriorates.
A cellular glass or cellular plastic type of insulation is fast becoming accepted as an ideal insulation.
Its cellular structure provides exceptionally high
resistance to water and water vapor. The cellular
glass, being inorganic, is fire-proof. Cellular plastic
which is also available is self-extinguishing.
When located out of doors, insulation must be
weatherprooFed unless, of course, it is inherently
waterproof.
.
3-81
CHAPTER 4. STEAM PIPING
This chapter describes practical design and layout
techniques for steam piping systems. Steam piping
differs from other systems because it usually carries
three fluids - steam, water and air. For this reason,
steam piping design and layout require special consideration.
c JERAL SYSTEM DESIGN
Steam systems are classified according to piping
arrangement, pressure conditions, and method of returning condensate to the boiler. These classifications are discussed in the following paragraphs.
PIPING
ARRANGEMENT
A one- or two-pipe arrangement is standard for
steam piping. The one-pipe system uses a single pipe
to supply steam and to return condensate. Ordinarily, there is one connection at the heating unit
for both supply and return. Some units have two
connections which are used as supply and return
connections to the common pipe.
A two-pipe steam system is more commonly used
in air conditioning, heating, and ventilating applications. This system has one pipe to carry the steam
supply and another to return condensate. In a two‘: system, the heating units have separate conn.-ctions for supply and return.
The piping arrangements are further classified
with respect to condensate return connections to the
boiler and direction of flow in the risers:
1. Condensate return to boiler
a. Dry-return - condensate enters boiler above
water line.
b. Wet-return - condensate enters boiler below water line.
2. Steam flow in riser
a. Up-feed - steam flows up riser.
b. Down-feed - steam flows down riser.
PiESSURE CONDITIONS
Steam piping systems are normally divided into
five classifications - high pressure, medium pressure,
low pressure, vapor and vacuum systems. Pressure
ranges for the five systems are:
1. High pressure - 100 psig and above
2. Medium pressure - 15 to 100 psig
3. Low pressure - 0 to 15 psig
4. Vapor - vacuum to 15 psig
5. Vacuum - vacuum to 15 psig
Vapor and vacuum systems are identical except
that a vapor system does not have a vacuum pump,
but a vacuum system does.
CONDENSATE RETURN
The type of condensate return piping from the
heating units to the boiler further identifies the
steam piping system. Twoarrangements, gravity
and mechanical return, are in common use.
When all the units are located above the boiler
or condensate receiver water line, the system is described as a gravity return since the condensate returns to the boiler by gravity.
If traps or pumps are used to aid the return of
condensate to the boiler, the system is classified as a
mechanical return system. The vacuum return
pump, condensate return pump and boiler return
trap are devices used for mechanically returning
condensate to the boiler.
CODES AND REGULATIONS
All applicable codes and regulations should be
checked to determine acceptable piping practice for
the pxticular application. These codes usually dictate piping design, limit the steam pressure, or
qualify the selection of equipment.
WATER CONDITIONING
The formation of scale and sludge deposits on the
boiler heating surfaces creates a problem in generating steam. Scale formation is intensified since
scale-forming salts increase with an increase in temperature.
Water conditioning in a steam generating system
should be under the supervision of a specialist.
PAR-I.
3-82
3. PIPING DESIGN
L
TABLE 25-RECOMMENDED HANGER
SPACINGS FOR STEEL PIPE
DISTANCE BETWEEN SUPPORTS (FT)
NOM.
PIPE SIZE
(in.1
Average Gradient
1” in 10’
J/i
I/?” in 10’
’
1
1%
1%
9
13
16
19
2
21
I
;
6
I
;
10
14
17
i;
19
40
33
‘A” in 10’
I
I
1
5
8
13
23
I
25
NOTE: Data is bared on standard wall pipe filled with water and
overage number of fittings.
Courtcsv or Crnne co.
PIPING
ONE-PIPE SYSTEM
SUPPORTS
All steam piping is pitched to facilitate the flow
of condensate. Table -3.5 contains the recommended
support spacing for piping pitched for different gradients. The data is based on Schedule 40 pipe filled
with water, and an average amount of valves and
fittings.
PIPING DESIGN
A steam system operating for air conditioning
comfort conditions must distribute steam at all operating loads. These loads can be in excess of design
load, such as early morning warmup, and at extreme
1’K;.
(30 - &JK-I’Il’1.:,
partial load, when only a minimum of heat is necessary. The pipe size to transmit the steam for a design load depends on the following:
I. The initial operating pressure and the allowable pressure drop thru the system.
2. The total equivalent length of pipe in the
longest run.
3. Whether the condensate flows in the same
direction as the steam or in the opposite direction.
The major steam piping systems used in air-conditioning applications are classified by a combination of piping arrangement and pressure conditions
as follows:
1. Two-pipe high pressure
2. Two-pipe medium pressure
3. Two-pipe low pressure
4. Two-pipe vapor
5. Two-pipe vacuum
6. One-pipe low pressure
IJl’F1’.1:11
(;KAVITY SYSTEM
A one-pipe gravity system is primarily used on
residences and small commercial establishments.
Fig. 90 shows a one-pipe, upfeed gravity system. The
steam supply main rises from the boiler’to a high
point and is pitched downward from this point
around the extremities of the basement. It is normally run full size to the last take-off and is then
reduced in size after it drops down below the boiler
water line. This arrangement is called a wet return.
If the return main is above the boiler water line,
it is called a dry return. Automatic air vents are required at all high points in the system to remove
non-condensable gases. In systems that require long
mains, it is necessary to check the pressure drop and
make sure the last heating unit is sufficiently above
the water line to prevent water backing up from. the
boiler and flooding the main. During operation,
steam and condensate flow in the same direction in
the mains, and in opposite direction in branches and
risers. This system requires larger pipe and valves
than any other system.
The one-pipe gravity system can also be designed
as shown in Fig. 91, with each riser dripped separately. This is frequently done on more extensive
systems.
Another type of one-pipe gravity system is the
down-feed arrangement shown in Fig. 92. Steam
flows in the main riser from the boiler to the building attic and is then distributed throughout the
building.
FIG.
I) 1 - ONI:-PII’IC
GRAVITY
DRIPPED
S~sTl:;cf
WITH
I<ISI-KS
TWO-PIPE SYSTEM
A two-pipe gravity system is shown in Fig. 93. This
system is used with indirect radiation. The addition
of a thermostatic valve at each heating unit adapts
it to a vapor or a mechanical vacuum system. A gravity system has each radiator separately sealed by drip
loops on a dry return or by. dropping directly into
a wet return main. All drips, reliefs and return risers
from the steam to the return side of the system must
be sealed by traps or water loops to insure satisfactory operation.
If the air vent on the heating unit is omitted, and
the air is vented thru the return line and a vented
condensate receiver, a vapor system as illustrated in
Fig. 97’ results.
The addition of a vacuum pump to a vapor system classifies the system as a mechanical vacuum
em. This arrangement is shown in Fig. 95.
CHART 26-PIPE SIZING+
5 6
810
2 0
4 0 60
100
200
400
1000
600
2000
4 0 0 0 1 0 0 0 0
6000
2 0000
50000
100 0 0 0
S T E A M F L O W RATE-LB/HR
*Use Chart 27 to determine steam velocity at initial saturated steam pressures other than 0 psig.
Charts
26 and 27 from Heating
Ventilating Air Conditioning Guide J 959. Used by permission.
(:~1.\l”I‘l:l<
I. s’I‘l-.\,\l l’ll’lN(;
3-85
CHART 27-VELOCITY CONVERSION*
RECOMMENDATIONS
The following recommcnd~ttioris are for use when
sizing pipe for the various systems:
Two-Pipe High Pressure System
60000
40
000
3 0 0 0 0
20000
10000
0000
6000
u
-
1000
0 0 0
G
2
e
600
600
-
4 0 0
4 0 0
300
300
1000
000
200
100 40
0
5 IO 2 0
100
140
100 “- 2 0 0
PkSSURE-PSIG
60RO
~~__
SATURATED
*See Exomde
STEAM
3, page 89, for use of chart
PIPE SIZING
GENERAL
Charts and tables have been developed which are
used to select the proper pipe to carry the required
steam rate at various pressures.
,hart -36 is a universal chart for steam pressure of
0 to 200 psig and for a steam rate of from 5 to 100,000
pounds per hour. However, the velocity as read
from the chart is based on a steam pressure of 0 psig
and must be corrected for the desired pressure from
Chart 27. The complete chart is based on the Moody
friction factor and is valid where condensate and
steam flow in the same direction.
Tables 26 tl~~ 31 are used for c1uic.k selection at
specific steam pressures. Clrtr~t _“(, has been used to
tabulate the capacities shown in Tnb1r.r 2~; thm -38.
The capacities in TnD1e.s 23 t/cl-u 3~ are the results
of tests conducted in the ASHAE laboratories. Suggested limitations for the USC of these tables are
shown as notes on each table. In addition, Table 31
shows the total pressure drop I’or two-pipe low pressure steam systems.
This system is used mostly in plants and occasionally in commercial installations.
1. Size supply main and riser for a maximum drop
of 25-30 psi.
2. Size supply main and risers for a maximum
friction rate of 2-10 psi per 100 ft of equivalent
pipe.
3. Size return main and riser for a maximum pressure drop of 20 psi.
4. Size return main and riser for a maximum friction rate of 2 psi per 100 ft of equivalent pipe.
5. Pitch supply mains G in. per 10 ft away from
boiler.
6. Pitch return mains ti in. per 10 ft toward the
L
boiler.
7. Size pipe from Table 26.
Two-Pipe Medium Pressure System
This system is used mostly in plants and occasionally in commercial installations.
1. Size supply main and riser for a maximum
pressure drop of 5-10 psi.
2. Size supply mains and risers for a maximum
friction rate of 2 psi per 100 ft of equivalent
pipe.
3. Size return main and riser for a maximum pressure drop of 5 psi.
4. Size return main and riser for a maximum
friction rate of 1 psi per 100 ft of equivalent
pipe.
5. Pitch supply mains ‘/4 in. per 10 ft away from
the boiler.
6. Pitch return mains ‘/4 in. per 10 ft toward the
boiler.
7. Size pipe from Table 27.
Two-Pipe Low Pressure System
This system is used for commercial, air conditioning, heating and ventilating installations.
1. Size supply main and risers for a maximum
pressure drop determined from Table 31, depending on the initial system pressure.
2. Size supply main and riser for a maximum friction rate of 2 psi per 100 ft of equivalent pipe.
3. Size return main and riser for a maximum
TABLE 26-HIGH PRESSURE SYSTEM PIPE CAPACITIES (150 psig)
Pounds Per Hour
PRESSURE DROP PER 100 FT
PIPE SIZE
(in.)
l/s psi
‘Vi psi (4
(2 02)
‘h psi (8 ox) S/4 psi
02)
psi
(12 o z ) 1
SUPPLY MAINS AND RISERS
=/
I
29
58
130
203
412
683
1,237
1,855
2.625
41858
7,960
16,590
30,820
48,600
1.
1%
1%
2
2%
3
3%
4
5
6
8
10
12
1
41
I
583
959
1,750
2,626
3.718
6;875
11,275
23,475
43,430
68,750
RETURN
1
156
313
%
1%
1 ‘h
2
2%
3
3%
4
5
6
,
650
1.070
21160
3,600
6,500
9,600
13,700
25,600
42,000
58
825
1,359
2,476
3,715
5.260
9;725
15,950
33,200
61,700
97,250
1,167
1,920
3,500
5,250
7,430
13;750
22,550
46,950
77,250
123,000
M A I N S AND R I S E R S
232
462
960
1.580
3;300
5,350
9,600
14,400
20,500
38,100
62,500
82
I
360
690
1,500
2,460
41950
8,200
15,000
22,300
31,600
58,500
96,000
465
910
1
1,950
3,160
6,400
10,700
19,500
28,700
40,500
76,000
125,000
pressure drop detcrminetl from Tlrble 31, dcpending on the initial system pressure.
-1. Size return main and riser for a maximum friction rate of I/? psi per 100 ft of equivalent pipe.
5. Pitch mains I/~ in. per 10 ft away from the
boiler.
F. Pitch return mains Q$ in. per 10 Et toward the
boiler.
2 psi (32 01)
(16 0x1
130 - 180 de--Max
1
116
233
523
813
1,650
2,430
4,210
6,020
8.400
15;ooo
25,200
50,000
90,000
155,000
/
10 psi
5 psi
Error 8%
184
369
827
1.230
2;ooo
3,300
6,000
8,500
12.300
2lj200
36,500
70,200
130,000
200,000
I
300
550
1,230
1.730
3;410
5,200
9,400
13,100
19.200
331100
56,500
120,000
210,000
320,000
I
420
790
1,720
2,600
4,820
7,600
13,500
20,000
28.000
471500
80,000
170,000
300,000
470,000
1 - 2 0 prig - Max Return Pressure
560
1,120
890
1,780
2,330
3,800
7,700
12,800
23,300
34,500
49,200
91,500
150,000
3,700
6,100
12;300
20,400
37,200
55,000
78,500
146,000
238,000
6. Pitch return mains ‘/, in. per 10 It towd
boiler.
:
the
7. Size pipe from Tables -78 th?l 30.
Two-Pipe Vacuum System
This system is used in commercial installations.
Two-Pipe Vapor System
1. Size supply main and riser for a masimum
pressure drop of !h - 1 psi.
2. Size supply main and riser for a maximum friction rate of !h - I/? psi per 100 ft of equivalent
pipe.
This system is used in commercial and residential
installations.
3. Size return main and riser for a maximum pressure drop of I$$ - 1 psi.
1. Size supply main and riser for a maximum
pressure drop of ‘,/1,; - t/s psi.
4. Size return main and riser for a maximum friction rate of Q$ - I/? psi per 100 ft of equivalent
pipe.
2. Size supply main and riser for n maximum
friction rate of !/lr; - IA3 psi per 100 It of equivalent pipe.
5. Pitch supply mains !/( in. per 10 ft away from
the boiler.
.?. Size return main and supply for a m;1ximum
pressure drop of \il; - xY psi.
4. SiLe return main and supply for ;t maximum
friction raw of )i,; - I$(? psi per 100 ft of equivalent pipe.
One-Pipe Low Pressure System
5. Pitcil supply 51 i n . per 10 ft away
Iwiler.
This system is usctl o n sm;dl commercial and
rcsitlential systems.
front t h e
G. Pitch return mains !/r in. per 10 Et townrcl
boiler.
7. Sire pipe from Tables 2S tht,~~ 30.
the
TABLE 27-MEDIUM PRESSURE SYSTEM PIPE CAPACITIES (30 psig)
Pounds Per Hour
PIPE SIZE
(in.)
PRESSURE DROP PER 100 FT
‘/s psi (2 02)
% psi (4 02)
=/4
1
1 ‘A
1%
2
2%
3
3%
4
5
6
8
10
12
15
31
69
107
217
358
651
979
1,386
2,560
4,210
8,750
16,250
25,640
?h
1
1%
1%
2
115
230
485
790
1,575
2%
3
3%
4
5
2,650
4,850
7,200
10,200
19,000
3,900
7,100
10,550
15,000
27,750
6
3 1,000
45,500
RETURN
.
‘ii psi (8 02)
9% psi (12 02)
MAINS
AND
170
340
710
1,155
2,355
2
1 psi (16 oz)
2 psi (32 OL)
25 - 35 psig - Max Error 8%
SUPPLY MAINS AND RISERS
22
31
46
63
100
141
154
219
313
444
516
730
940
1,330
1,414
2,000
2,000
2,830
3,642
5,225
6,030
8,590
12,640
17,860
23,450
33,200
36,930
52,320
38
77
172
267
543
924
1,628
2,447
3,464
6,402
10,240
21.065
40,625
64,050
45
89
199
309
627
1,033
1,880
2,825
4,000
7,390
12,140
25,250
46,900
74,000
63
125
281
437
886
1,460
2,660
4,000
5,660
10,460
17,180
35,100
66,350
104,500
0 - 4 psig - Max Return Pressure
RISERS
245
490
1,025
1,670
3,400
308
615
1,285
2,100
4,300
365
730
1,530
2,500
5,050
5,600
10,250
15,250
1,600
40,250
7,100
12,850
19,150
27,000
55,500
8,400
15,300
22,750
32,250
60,000
65,500
83,000
,98,000
TABLE 2B-LOW PRESSURE SYSTEM PIPE CAPACITIES
Pounds Per Hour
CONDENSATE
NOM.
PIPE
SIZE
(in.)
vi
1
1%
1%
2
2%
3
3 ‘h
4
'A6 psi (1 02)
3.5
9
17
36
56
108
174
318
462
726
12
‘/a psi (2 02)
1 3.5
11 1
/
12
FLOWING
WITH
THE
FLOW
PRESSURE DROP PER 100 FT
% psi (12 02)
% psi (4 02)
‘/a psi (8 02)
S A T U R A T E D P R E S S U R E (PSIG)
1 3.5
1 12 1 3.5
I 12 I 3.5 I 12 I
14 /
161
20 1
241
29
54
31
37
46
66
70
96
111
100
120
234
194
378
310
550
660
800
990
1,160
1,410
2,100 2 , 4 4 0
3,350 3 , 9 6 0
7,000
8,100
12,600
15,000
3 , 7 0 0 ( 1 6 , 5 0 0 ) 1 9 , 5 0 0 1 23;400 1 28;400 / 3 3 , 0 0 0
The weight-flow rater at 3.5 psig cctn be used +CJ cover
8 to 16 psig with an error not exceeding 8 percent.
STEAM
so+. press.
35
66
138
210
410
660
1,160
1,700
2,400
4,250
7,000
14,300
26,000
40.000
1 psi
3.5
I
2 psi
12
50
95
200
304
590
950
1,670
2,420
3,460
6,100
10,000
20,500
37,000
I
3.5
60
114
232
360
710
A
1,150
1,950
2,950
4,200
7.500
A
11,900
24,000
42,700
67,800
from 1 to 6 prig, and the rater at 12 prig con be used to cover sat. press.
Tobk 26 thru 28 from Heating
Venfilafing
I
12
”
73
137
280
430
850
1,370
2,400
3,450
4,900
8.600
A
14,200
29,500
52,000.
81,000’
from
Air Conditioning Guide 1959. Used by permission.
TABLE 29-RETURN MAIN AND RISER CAPACITIES FOR LOW PRESSURE STEAM SYSTEMS
PIPE
SIZE
‘5% psi (l/z 02)
(in.) W e t * 1 D r y / VclC
PRESSURE DROP PER 100 FT
l/24 psi (75 02)
wet*
Vat
DV
‘A6 psi (1 01)
Wet* 1 Dry Vat
‘Al psi (2 02)
Wet+ / Dry
VlX
RETURN
1
125
213
-.338
700
62
1.-30
206
470
145
248
393
810
71
149
236
535
42
143
244
388
815
1.180
760
1,580
868
1,360
J/4
1-,.
‘A
1%
2
2%
3 % 21' .7 580 8 0
4
3:880
5
6
1
1'970 1460
2:930
175
300
475
1,000
80
168
265
575
1,680
?4
48
48
143
1
113
113
244
1 %
240
248
388
1%
375
375
815
2
750
750
1,360
2%
2,180
3
3,250
3%
4,480
4
7,880
5
12,600
* V a t values may be u s e d f o r w e t r e t u r n r i s e r s a n d m a i n s .
7. Size supply main and dripped
40
113
248
375
750
runouts from
200
350
600
950
2,000
283
494
848
1,340
2,830
2,350 1 1,2301 2,380
3,350
1,360
3,350
4,730
71; 5 6 0 1 , 3 0 0
15,500
27,300
143,800
RISERS
48
113
248
375
750
249
426
674
1,420
2,380
3,800
5,680
7,810
13,700
22,000
Vctltilating
Air Conditioning:
48
350
113
600
248
950
375 2,000
7501 3.3501
I 5.3501
8,000
11,000
19,400
31,000
I
I
C;uidc, 19’19. Used by
494
848
1,340
2,830
I 41730
1 7.560
'
,~
11,300
15,500
27,300
43,800
permission.
8. Size undripped runouts from Table 30, Column I;.
9. Size upfeed risers from Table 30, Cqlunn D.
10. Size downfeed supply risers from Table 28.
il. Pitch supply mains vi in. per 10 ft away from
boiler.
Pounds Per Hour
CONDENSATE FLOWING AGAINST STEAM FLOW
I TWO-PIPE
’
SYSTEM
Vertical
Horizontal
B*
ct
8
1 ONE-PIPE
12. Pitch return mains M in. per 10 ft toward the
boiler.
SYSTEM
Up-feed Vertical
Supply
CO”Risers
nectars
Riser
RUllouts
DIt
E
F
14
31
48
97
-
1
1 'h
1%
2
9
i9
27
49
6
11
20
38
72
7
16
23
42
7
7
16
16
23
2'h
3
3%
4
5
159
282
387
511
1,050
99
175
288
425
788
116
200
286
380
-
-
42
65
119
186
278
1,800
3,750
7,000
1 1,500
22,000
1,400
3,000
5,700
9,500
19,000
-
-
545
-
6
8
10
12
16
115
241
378
825
I:ro~n Iqe:lting
TABLE 30-LOW PRESSURE SYSTEM
PIPE CAPACITIES
%
350
600
950
2,000
175
300
475
1,000
1,680
2,680
4,000
5,500
9,680
15,500
ToDie 23.
A
142
103
249
217
426
3401
674
740 1 1,420
250
425
675
1,400
2 , 1 3 0 3 , 3 0 02 , 2 0 0 1 , 5 6 03 , 2 5 0 2 , 1 8 04 , 0 0 0 2 , 6/
/2
8 0 , 5 0 0 1 , 7 5 04 , 0 0 0 2 , 6 8 05 , 5 0 0 3 , 7 5 02 , 2 5 0 3 , 2 3 03 , 8 0 0 5 , 6 8 05 , 3 5 0 8 , 0 0 02 , 5 0 0 3 , 5 8 05 , 3 5 0 8 , 0 0 0
4,580 3,350 4,500 5,500 3,750 5,500 7,750 4,830 7,810 11,000 5,380 11,000
7,880
9,680
13,700
19,400
Il2,6001
/1 5 , 5 0 0 1
/22,000 1
~31,000~
I
RETURN
PIPE
SIZE
(in.)
‘/a psi (8 02)
wet*
VfVJ
W
MAINS
100
175
300
475
, 1,000
1,680 1 9501
‘/i psi (4 02)
Wet* Dry
VW2
*DO not use Column B for pressure drops less than %6 psi/100ft of
e q u i v a l e n tl e n g t h . U s eC h a r t 2 6 , page 8 4 .
i P i t c h of h o rtu o n oIr u n o v t rt o r i s e r s h o u l d b e n o t l e s s t h a n ‘ % i n . / f t .
W h e r e t h i s p i t c h c a n n o t b e o b t a i n e d , r u n o u tosv e r B f t i n l e n g t h s h o u l d
be one pipe size larger than called for in this table.
$Do not use Column D for pressure drops less than l/24 psi/100ft of
equivalent run except on sizes 3 in. and larger. Use Chart 26, page 84.
Fmrn Hcatina Ventilntina Air Conditioning Guide,
Used by permission.
1959.
Use of Table 31
Example I - Determine Pressure Drop for Sizing Supply *
and Return Piping
Given:
Two-pipe low pressure steam system
Initial steam pressure - 15 psig
Approximate supply piping equivalent length - 500 ft
.Approximate return piping equivalent length - 500 ft
Find:
1. Pressure drop to size supply piping
2. Pressure drop to size return piping
Solution:
1. Refer to TnOle 31 for an initial steam pressure of 15
psig. The total pressure drop should not exceed 3.75 psi
in the supply pipe. Therefore, the supply piping is sized
for a total pressure drop of 3.75 or 3/4 psi per 100 ft of
equivalent pipe.
2. .\lthough ‘v, psi is intlicatctl in .Strj> 1, Item 4 under the
two-pipe low pressure system recommends a maximum
of I/ psi for return piping. Therefore, use I/* psi per 100
ft of equivalent pipe.
’
TABLE 31-TOTAL PRESSURE DROP FOR
TWO-PIPE LOW PRESSURE STEAM PIPING
SYSTEMS
Friction Rate
fixainple
2 illustrates the method used to tleterlriction rate for sizing pipe when the total
system pressure drop recommendation (supply prcssure drop plus return pressure drop) is known and
the approximate equivalent length is known.
lnirie tlw
( p s i )
(prig)
Example
2
2
5
- Determine Friction Rate
Civcn:
Four systems
liquivnlent length of each system - 400 ft
Total pressure drop of systems - I/?, :v,, I, and 2 psi
1%
2%
3 vi
10
I5
‘/a
1‘/4
2 ‘/!I
3%
Fintl:
I. Size of largest pipe not exceeding tlcsign friction rate.
2. Swam velocity in pipe.
Find:
Friction rate for each system
Solution:
SYSTEhl
NUMl3ER
(psi)
1%
SYSTEM
EQUIV.
LENGTH
TOTAL
SYSTEXI
PRESS.
DROP
(4
(Psi)
Solution:
1. Enter Idtorn of Clrnrt 26 at 67.50 Ib/hr anti proceed vutically to the 100 psig line (dotted line in Clinrt -16). Then
move ol,liquely to the 0 psig line. From this point proceed vertically up the chart to the smallest pipe size not
csccetling 2 psi per 100 ft of equivalent pipe and read
S!/, in.
FRICTION
R,\TE
FOR PIPE
SIZING
(per 100 ft)
400
(400/ 100) (x) = %
2. The velocity of steam at 0 psig ~1s read from Clrnrt -76 is
16,000 fpm. Enter the left side of C/r& 2i at 16,000 fpm.
I’roccctl ol~liquely downward to the 100 psig line andhorizontally across to the right sitlc of the chart (dotted
lint in C/la,l 37). The velocity at 100 psig is 6100 fpm.
x = I/*
400
(400/ 100) (x) = :fi
400
(400/ 100) (x) = 1
400
(400/ 100) (x) = 2
x =siti
x =!A
Exnnzple
f illustrates a design problem for sizing
pipe on a low pressure, vacuum return system.
x = 1/x
Example 4 - Sizing Pipe for a Low Pressure, Vacuum
Use of Charts 26 and 27
Example
3
-Determine
Velocity
Return System
Steam
Supply
Main
and
Final
Given:
Six units
Steam requirement per unit - 72 lb/hr
Layout as illustrated in
thru 98
Threaded pipe and fittings
Low pressure system - 2 psi
Given:
Friction rate - 2 psi per 100 ft of equivalent pipe
Initial steam pressure - 100 psig
Flow rate - 6750 lb/hr
72
72
72
72
Frc.96 - Low PRE~SURESTEAM
72
SUPPLY MAIN
72
3-90
I’r\K’l‘ 3 . 1’ll’lNG
I’rcssurc drop for the supply main is equal
lent lcngttl times pressure tlrop per 100 It:
DESIGN
to the equiva-
167.5 X .25/ IO0 = .42 psi
This is within the rccommcntfctl
(I psi) for the sl;pply.
The I)ranch connection for I;ig.
ner at the same friction rate.
6’
maximum
pressure
drop
97 is sired in a similar man-
From T&/r 30 the hori/.ontal runout pipe size for a load
of i2 II) is 2l,/2 in. ant1 the vertical riser size is 2 in.
Convert all the fittings to equivalent pipe lengths. and add
to the actual pipe length.
Equivalent
F IG.
97
-
Low
RUNOUT
PKESSURE
Find:
Size of pipe and total pressure
Note: Total pressure drop
exceed one-half the initial
drop is required for quiet
AND
RISER
drop
in the system should
pressure. A reasonahly
operation.
pipe
lengths
I - 2th in. 4.5” cl1
1 - 21/ in. 90” ell
I - 2 in. 90” ell
I - 2 in. gate valve
Actual pipe length
Total
never
small
Solution:
Determine the design friction rate hy totaling the pipe
length and adding 50% of the length for fittings:
equivalent
3.2
4.1
3.3
2.3
11.0
length
23.9 ft.
I’rcssure drop for branch runout
23.9 X .23/lOO
.
and riser is
= ,060 psi
The vacuum return main is sized from Tab/e 29 by starting
at the last unit “G” and adding each additional load hetween unit “G” and the boiler.
Each riser - 52 lb per hr, s/11 in.
II5 + 11 + 133 = 259
259x .50=
130
389 ft equiv length
SECTION
Check pipe sizing recommendations for maximum friction
rate from “Two-Pipe Vacuum System,” Ztem 2, l/8-1/? psi.
Check Tab/e 31 to determine recommended maximum pressure drop for the supply and return mains (I/? psi for each).
STEAM
LOAD
PIPE SIZE
(in.)
Oh/W
Design friction rate = l/3.89 X (l/2 + l/2) = l/4 psi per 100
ft. The supply main is sized hy starting at the last unit “G”
and adding each additional load from unit “G” to the
boiler; use Table 28. The following tabulation results:
SECTION
STEAM
LOAD
P/W
F-G
E-F
D-E
C -D
IS-C
A - n
72
144
216
288
360
432
PIPE SITE
(in.)
1%
2
2
2’/
3
3
Convert the supply main fittings to equivalent lengths of
pipe and add to the actual pipe length, Table II, page 17.
Equivalent
pipe
lengths
1 - 1% in. side outlet tee
2 - 1% in. ells
1 - 2 in. reducing tee
i - 2 in. run of tee
2 - 2 in. ells
1 - 21/~ in. reducing tee
I - 3 in. reducing tee
2 - 3 in. ells
1 - 3 in. run of tee
Actual pipe length
Total
equivalent
length
7.0
4.6
4.7
3.3
6.6
5.6
7.0
15.0
5.0
115.0
167.5 ft.
Convert the return main fittings to equivalent pipe lengths
and add to the actual pipe length, Table 11, page 17.
’ Equivalent pipe lengths
1 _ 3h in. run of tee
5 - % in. 90” ells
1 - I in. reducing tee
1 - 1 in. run of tee
2 - 1 in. 90” ells
1 - II/~ in. reducing tee
3 - I IA in. 90” ells
1 - 1% in. run of tee
.Actual pipe length
Total equivalent length
1.4
7.0
2.3
1.7
3.4
3.1
6.9
2.3
133.0
161.1 ft.
Pressure drop for the return equals
161.1 X .25/100 = ,404 psi
Total return pressure drop is satisfactory since it is within
the recommended maximum pressure drop (t/B - 1 psi) listed
in the two-pipe vacuum return system. The total system
pressure drop is equal to
,420 + .060 + ,404 = ,884 psi
.
This total system pressure drop is within the maximum 2 psi
recommended (1 psi for supply and 1 psi for return).
(:I-I.\l’~I‘El<
4.
S’I‘E.\XI
3-91
I’II’ING
72
72
FIG.
98 -Low
72
PRESSURE
‘ING APPLICATION
The use and selection of steam traps, and condensate and vacuum return pumps are presented
in this section.
Also, various steam piping diagrams are illustratcd to familiarize the engineer with accepted
piping practice.
STEAM TRAP SELECTION
The primary function of a steam trap is to hold
steam in a heating apparatus or piping system and
allow condensate and air to pass. The steam remains
trapped until it gives up its latent heat and changes
to condensate. The steam trap size depends on the
following:
1. Amount of condensate to be handled by the
trap, Ib/hr.
2. Pressure differential between inlet and discharge at the trap.
3. Safety factor used to select the trap.
Amount of Condensate
The amount of condensate depends on whether
the trap is used for steam mains or risers, or for the
heating apparatus.
The selection of the trap for the steam mains or
risers is dependent on the pipe warm-up load and
the radiation load from the pipe. Warm-up load is
the condensate which is formed by heating the pipe
surface when the steam is first turned on. For practical purposes the final temperature of the pipe is
the steam temperature. Warm-up load is determined from the following equation:
c = ‘V x (I/.- ti) x .114
I
A, x T
72
72
VACUUM
72
RETURN
where:
C, = Warm-up condensate, lb/hr
W = Total weight of pipe, lb (Tables 2 and 3,
pages -7 and 3)
I~ = Final pipe temperature, F (steam temp)
ti = Initial pipe temperature, F (usually room
tern p)
.114 = Specific heat constant for wrought iron or
steel pipe (.092 for copper tubing)
h, = Latent heat of steam, Ktu/lb (from steam
tables)
T = Time for warm-up, hr
The radiation load is the condensate formed by
unavoidable radiation loss from a bare pipe. This
load is determined from the following equation and
is based on still air surrounding the steam main or
riser:
c = L X I\ X (tf - ti)
2
f/l
where:
C, = Radiation condensate, lb/hr
L = Linear length of pipe, ft
K = Heat transmission coefficient, Btu/(hr)
(linear Et) (deg F diff between pipe and
surrounding air), Table jf, Part I
t,, ti, II, explained previously
The radiation load builds up as the warm-up
load drops off under normal operating conditions.
The peak occurs at the mid-point of the warm-up
cycle. Therefore, one-half of the radiation load is
added ‘to the warm-up load to determine the amount
of condensate that the trap handles.
3-a
PAR?‘
3. PlPlNG D E S I G N
L
Pressure
Differential
T!le Ilressure tlilfercntial. across the trap is determined at dcsigri conditions. IL‘ a vacu~~m exists on
the tliscllarge side of the trap, the vacuuln is added to
the inlet side pressure to determine the differential.
Safety Factor
Good design practice dictates the use of safety
factors in steam trap selection. Salety factors from
2 to 1 to as high as 8 to 1 may be requirctl, and for
the following reasons:
1. The steam pressure at the trap inlet or the
back prcssurc
at the trap discharge may vary.
This changes the steam trap capacity.
2. If the trap is sized Eor normal operating load,
condensate may back up into the steam lines
or apparatus during start-up or warm-up opera(
tion.
3. II the steam trap is selected to discharge a full
and continuous stream of water, the air could
not be vented from the system.
The following guide is used to determine the
safety factor:
DESIGN
Draining steam main
Draining steam riser
SAFETY
Before reducing valve
Before shut-off valve
(closed part of time)
Draining coils
Draining
apparatus
3 to 1
3 to 1
3 to 1
5
-
Steam
Trap
Selection
for
Solution:
I. The warm-up
equation:
Dripping
Main to Return Line
Given:
Steam main - 10 in. diam steel pilje, 50 ft long
Steam pressure - 5 psig (227 F)
Root11 temperature
- 70 F tli) (stealn main in space)
Warm-up time - 15 minutes
Steam trap to drip main into vacuum rettirn line
(2 in. vacuum gage design)
load is determined from the following
iv x pr - ti) x .I I4
c, =
llL
x T
where: 1%‘~ 40.48 lb/ft X 50 ft (Tnble 2)
tf= 227 F
ti= 7OF
hL = 960 Btu/lb (from
steam tables)
T = .25 hr
c, =
2024 x (227 - TO) x ,114
960 x .?5
= 156 lb/hr of condensate
2. The radiation load is calculated by using the following
equation:
L x I\’ x (I, - li)
c, =
‘11
where: L = 50 ft
K = 6.41 Btu/(hr) (linear foot)
(deg F diff between pipe and air)
from Table 5-I, Part I
tf = 227 F
ti
= 70 F
tzL = 960 Rtu/ll, (from
steam taD1r.s)
c =50 X 6.41 x (227- 70)
2
960
= 52 lb/hr‘of condensate
3.
When the steam trap is to be used in a high pressure system, determine whether or not the system
is to operate und& low pressure conditions at cerain intervals such as night time or weekends. If this
.-condition is likely to occur, then an additional
safety factor should be considered to account for
the lower pressure drop available during night time
operation.
Example 5 illustrates the three concepts mentioned previously in trap selection-condensate handled, pressure differential and safety factor.
Example
Warm-up load.
Radiation load.
Total condensate load.
Specifications for steam trap at end of supply main.
FACTOR
3 to 1
2 to 1
2 to 1
3 to 1
Between boiler and end of main
Find:
1.
2.
3.
4.
Supply
The
equal
load.
total condensate load for steam trap selection is
to the warm-up load plus one half the radiation
Total condensate load = C, + (I$$ X Cr)
= 150 + (I/? x 52)
= 176 lb/hr
4. Steam trap selection is dependent on three factors: condensate handled, safety factor applied to total condensate
load, and pressure differential across the steam trap.
The safety factor for a steam trap at the end of the main
is 3 to 1 from the table on this page. Applying the 3 to 1
safety factor to the total condensate load, the steam trap
would be specified to handle 3 X 176 or 528 Ib/hr of
condensate.
The pressure differential across the steam trap is determinecl I)y the pressure at the inlet and diskharge of the
steam trap.
Inlet to trap
= 5 psig
Discharge of trap = 2 in. vacuum (gage)
\\‘hen the discharge is under vacuum conditions, the
discharge vacuum is added to the inlet pressure for the
total pressure differential.
Pressure differential = 6 psi (approx)
Therefore the steam trap is selected for a differential
pressure of G psi and 528 Ib/hr of condensate.
.
s#-
AIR VENT
I-
VALVE ACCESS
PLUG-
THERMOSTATIC
DISC ELEMENT
AIR PASSAGE
INLET
VALVE AN0
ORIFICE
I
STEAM TRAP TYPES
The types of traps commonly used in steam systems arc:
Float
Flash
Y~ermostatic
Implllsc
_ loat & thermostatic
Lifting
Upright bucket
boiler return or
inverted bucket
alternating receiver
The description and use of these various traps
are presented in the following pages.
Float Trap
The discharge from the float trap is generally continuous, This type (Fig. 97) is used for draining condensate from steam headers, steam heating coils,
and other similar equipment. When a float trap is
used for draining a low pressure steam system, it
should be equipped with a thermostatic air vent.
Thermostatic Trap
The discharge from this type of trap is intermittent. Thermostatic traps are used to drain condensate from radiators, convcctors, steam heating
cnils, unit heaters and other similar equipment.
.iners are normally installed on the inlet side of
the steam trap to prevent dirt and pipe scale from
OUTLET
entering the trap. On traps used for radiators or
convectors, the strainer is usually omitted. I;ig. 100
shows a typical thermostatic trap of the bellows type
and I;ig. 101 illustrates a disc type thermostatic trap.
When a thermostatic trap is used for a heating
apparatus, at least 2 ft of pipe are provided ahead
of the trap to cool the condensate. This permits condensate to cool in the pipe rather than in the coil,
and thus maintains maximum coil efficiency.
Thermostatic traps are recommended for low
pressure systems up to a maximum of 15 psi. When
used in medium or high pressure systems, they must
be selected for the specific design temperature. In
addition, the system must be operated continuously
at that design temperature. Tlzis menns no night
setback.
Float and Thermostatic Trap
This type of trap is used to drain condensate from
blast heaters,,steam heating coils, unit heaters and
other apparatus. This combination trap (Fig. 102)
is used where there is a large volume of condensate
which would not permit proper operation of a thermostatic trap. Float and thermostatic traps are used
in low pressure heating systems up to a maximum
of 15 psi.
THERMOSTATIC
DISC ELEMENT\
ERMOSTATIC
INLET
I
OUTLET
FIG. IOO-THERMOSTATIC
TRAP , BELLOWSTYPE
L! FLOAT
L VALVE AND
ORfFlCE
Frc.102 -TYPIC.ALFLOATANUTHEKMOSTATICTRAP
I
OUTLET
INLET
BLOWOFF /
FIG. 105 - INVERXD I~UCKET TRAP WITH GUIDE
For medium and high pressure systems, the same
limitations as outlinctl for thermostatic traps apply.
‘Ipright Bucket Trap
The discharge of condensate from this trap (Fig.
is intermittent. h tlilfcrential pressure of at least
1 psi between the inlet and the outlet of the trap
is normally required to lift the condensate from the
bucket to the discharge connection. Upright bucket
traps are commonly used to drain condensate and
air from the blast coils, steam mains, unit heaters
and other equipment. This trap is well suited for
systems that have pulsating pressures.
103)
ing condensate from steam lines or equipment where
an abnormal amount of air is to be discharged and
where dirt may drain into the trap.
Flash Trap
The discharge from a Ilash trap (Fig. 106) is intermittent. This type of trap is used only if a pressure
differential of 5 psi or more exists between the steam
supply and condensate return. Flash traps may be
used with unit heaters, steam heating coils, steam
4
lines and other similar equipment.
Impulse Trap
Inverted Bucket Trap
The discharge from the inverted bucket trap (Figs.
104 and 105) is intermittent and requires a differential pressure between the inlet and discharge of
the trap to lift the condensate from the bottom of
the trap to the discharge connection.
Bucket traps are used for draining condensate and
air from blast coils, unit heaters and steam heating
-oils. Inverted bucket traps are well suited for drain-
t
Under normal loads the discharge from this trap
(Fig. IO/‘) is intermittent. When the load is heavy,
however, the discharge is continuous. This type of
trap may be used on any equipment where the pressure at the trap outlet does not exceed 25% of the
inlet pressure.
Lifting Trap
The lifting trap (Fig. 108) is an adaption of the
upright bucket trap. This trap can be used on all
steam heating systems up to 150 psig. There is an
OUTLET
ADJUSTABLE
ORIFICE
VALVE AN0
ORIFICE
OUTLET
AIR VENT
RE -EVAPORATING
CHAMBER
DRAIN PLUG
t
INLET
F1c.104 -IINVERTEDBUCKETTRAP
AND SLOWDOWN
VALVE OPENING
FIG. 106 -FLASH
TRAP
.
VENT TO
ATMOSPHERE-
FLOAT
INLET AND
t
Flc. 109 - UOIL.ER
OUTLET
RIXIJRN TRAP
OR
ALTERNAWNG
RECEIVER
FIG. 105 - 1xr~u1.s~ TRAP
of condensate return pumps are the rotary, screw,
turbine and reciprocating pump.
auxiliary inlet for high llressure steam, as illustrated
il- *he figure, to force the condensate to a point
Je the trap. This steam is normally at a higher
;r
pressure than the steam entering at the regular inlet.
Boiler Return Trap or Alternating Receiver
This type ol‘ trap is WA to return condensate to
a low pressure boiler. The boiler return trap (Fig.
fO9) does not hold steam as do other types, but
is an adaption of the lilting.trap. It is used in conjunction with a boiler to prevent flooding return
mains when excess pressure prevents condensate
from returning to the boiler by gravity. The boiler
trap collects condensate and equalizes the boiler and
trap pressure, enabling the condensate in the trap
to How back to the boiler by gravity.
CONDENSATE
RETURN
PUMP
Condensate return pumps are used for low
sure, gravity return heating systems. They are
mally of the motor driven centrifugal type and
2
:eiver and automatic float control. Other
HIGH PRESSURE
INLET
FIG. 108 - LIITIW
TRAP
presnorhave
types
The condensate receiver is sized to prevent large
Auctuations of the boiler water line. The storage
capacity of the reccivcr is approximately 1.5 times
the amount of condensate returned per minute, and
the condensate pump has a capacity of 2.5 to 3 times
normal flow. This relationship of pump and receiver
to the condensate takes peak condensation load into
account.
VACUUM
PUMP
Vacuum pumps are used on a system where the
returns are under a vacuum. The assembly consists
of a receiver, separating tank and automatic controls for discharging the condensate to the boiler.
Vacuum pumps are sized in the same manner as
condensate pumps for a delivery of 2.5 to 3 times the
design condensing rate.
PIPING LAYOUT
Each application has its own layout problem with
regard to the equipment location, interference with
structural members, steam condensate, steam trap
and drip locations. The following steam piping diagrams show the various principles involved. The
engineer must use judgment in applying these principles to the application.
Gate valves shown in the diagrams should be
used in either the open or closed position, ne?Je?’ fog’
tA~~ttlj?z~.
Angle and globe valves are recommended
lor throttling service.
In a one-pipe system gate valves are used since
they do not hinder the How of condensate. Angle
valves may be used when they do not restrict the
ffow of condensate.
All the figures show screwed fittings. Limitations
for other fittingi are described in Chnfiter 1.
+ STEAM MAIN
RISER
blRlP T O
CONDENSATE
RETURN
FIG. 1 IO - CONNECTION
TO
DRIITEI)R~srx
Steam Riser
Figz~es / 10 c1)7d I II illustrate steam supply risers
connected to mains with runouts. The runout in
Fig. 110 is connected to the bottom portion of the
main and is pitched toward the riser to permit condensate to drain from the main. This layout is used
only when the bser is dripped. If a dry return is
used, the riser is dripped thru a steam trap. If a
wet return is used, the trap is omitted.
Fig. 1 I I shows a piping diagram when the steam
riser is not dripped. In this instance the runout is
connected to the upper portion of the steam’ main
and is pitched to carry condensate from the riser to
the main.
Prevention of Water Hammer
If the steam main is pitched incorrectly when the
riser is not dripped, water hammer may occur as
illustrated in Fig. 112. Diagram “a” shows the runout partially filled with condensate but with
enough space for steam to pass. As the amount of
condensate increases and the space decreases, a wave
notion is started as illustrated in diagram “b”. As
the wave or slug of condensate is driven against the
turn in the pipe (diagram “c”), a hammer noise is
FIG. 112 -WATER HAMMER
caused. This pounding may be of sufficient force to
split pipe fittings and damage coils in the system.
The following precautions must be taken to prevent water hammer:
1. Pitch pipes properly.
.
2. Avoid undrained pockets.
3. Choose a pipe size that prevents high steam
velocity when condensate flows opposite to the
steam.
Runout Connection to Supply Main
Figure 113 illustrates two methods of connecting
runouts to the steam supply main. The method
using a 45” ell is somewhat better as it offers less
resistance to steam flow.
Expansion and Contraction
Where a riser is two or more floors in height, it
should be connected to the steam main as shown
RISER
(NOT DRlP,PED%D
//
STEAM
MAIN
FIG. 111 -
ACCEPTABLE
CONNECTION
TO
R I S E R (N O T D R I P P E D)
RECOMMENDED
FIG. 113 - RUNOUT CONNECTIONS
MOVEMENT
t
RAIN
AIR
LINE,
FIG. Ii4 - RISER CONNECTEI)TO ALLOW FOR
in Fig. 114. Point (A) is subject to a twisting movemcnt as the riser moves up and down.
Figure 115 shows a method of anchoring the
steam riser to allow for expansion and contraction,
Movement occurs at (A) and (B) when the riser
moves up and down.
HANDHOLE’
PLUG IN TEE
FOR CLEANOUT
FI G . 117 - RET~JRN
MA I N
LO O P
Obstructions
Steam supply mains may be looped over obstructions if a small pipe is run below the obstruction to
take care of condensate as illustrated in Fig. 116.
The reverse procedure is followed for condensate
return mains as illustrated in Fig. 117. The larger
pipe is carried under the obstruction.
Dripping
BOILER
REDUCING
WATER
COUPLING
Riser
A steam supply main may be dropped abruptly to
a ’ <:er level without dripping if the pitch is downWL . . . When the steam main is raised to a higher
level, it must be dripped a S illustrated in Fig. 318.
FIG. 118 - DRIPPING S TEAM MAIN
I‘his diagram shows the steam main dripped into a
wet return.
Figr~x I19 is one method of dripping a riser thru
RISER
STEAM TRAP
MOVEMENT
CONDENSATETE-RETURNMAIN
FI G . 115 - RISKR
AN C H O R
I’,\K
3-98
I
I. 3. I’ll’lh‘(;
,-CONTROL VALVE (NOTE
INVERTED LIFT
FITTING OR ELBOWS
NoTE ’ -.,\d
I)bSIGN
3)
r STRAINER
.
I
TOTAL LIFT
LIMIT 5 FT
II
f IN. PET COCK‘----h
A
WHEN END OF SUPPLY
MAIN,SEE FIG. 125
\-LIFT PIPE $“DIA.
OF VACUUM RETURN
15’ CHECK VALVE
FOR BREAKING VACUUM
%
I
VACUUM
RETURN
F R O
M
-NO JOINTS
LIFT PIPE
LIFT
IN
FITTING
DIRT LEG
15’ CHECK
FIG .
120 -
ONE-STEP CONDENSATE
Lggif+?? -y
LIFT
CONDENSATE
RETURN MAIN
a steam trap to a dry return. The runout to the return main is pitched toward the return main.
Vacuum
Lift
As described under vacuum systems, a lift is
sometimes employed to lift the condensate up to the
inlet of the vacuum pump. Figs. 120 and 122 show
a one-step and two-step lift respectively. The onestep lift is used for a maximum lift of 5 Et. For 5 t o 8
Et a two-step lift is required.
SOTES:
1. Flange or union is located to facilitate coil removal.
2. Flash trap may be used if pressure differential between steam and condensate return exceeds 5 psi.
3. When a bypass with control is required, see F&. 126.
4. Dirt leg may be replaced with a strainer. If so, tee
on drop can be replaced by a reducing ell.
5. The petcock is not necessary with a bucket trap or
any trap which has provision for passing air. The
great majority of high or medium pressure returns
end in hot wells or deaerators which vent the air.
FIG.
Steam
122
-HIGH
OR
MEDIUM
PRESSIJRE
COIL PIPING
Coils
Fig2ive.s 112 t/lrz~ 131 s h o w methods o f p i p i n g
THAN 2 STEPS
2ND STEP
I ST STEP
q
L/2
/iI
*MAXIMUM
LENGTH (A)
VACUUM
RETURN
/
FIG. 12 1 - TWO-STEP
CONDENSATE LIFT
-
steam coils in a high or low pressure or vacuum
steam piping system. The following general rules
are applicable to piping layout for steam coils used
in all systems:
1. Use full size coil outlets and return piping to
the steam trap.
2. Use thermostatic traps for venting only.
3. Use a 15’ check valve only where indicated on
the layout.
4. Six the steam control valve for the steam load,
not for the supply connection.
5. Provide coils with air vents as required, to
eliminate non-condensable gases.
6. Do not drip the steam supply mains into coil
sections.
.
,-CONTROL VALVE ( NOTE 2)
FLANGE OR UNION
CONTROL VALVE (NOTE 21
MAIN
‘-WHEN END OF SUPPLY
MAIN SEE FIG 125
AM SUPPLY MAIN
125 WHEN END
M SUPPLY MAIN
15’ CHECK VALVE
THERMOSTATIC
TRAP
FLOAT AND’
THERMOSTATIC TRAP
RETURN MAIN
NOTES:
1. Flange or union is located to facilitate coil removal.
2. When a bypass with control is required, see Fig. 126.
3. Check valve is necessary when more than one unit
is connected to the return line.
4. Dirt pocket is the same size as unit outlet. If dirt
pocket is replaced by a strainer, replace tee w i t h
a reducing ell from unit outlet to trap six.
CONDENSATE
RETURN MAIN
NOTES:
1. Flange or union is located to facilitate coil removal.
2. When a bypass with control is specified, see Fig. 126.
3. Check valve is necessary when more than one unit is
connected to the return line.
124 - VACUUM
FIG.
FIG. 123 - SINGLE
COIL
P
Y
Low
PRESSUKE
SY S T E M
STuni
COIL
I’II’IN(;
PIPING
GRAVITYRETURN
7. Do not pipe tempering coils and reheat coils
to a common steam trap.
8 . Multiple coils m a y be piped lo a common
steam trap if they have the same capacity and
the same pressure drop and if the supply is
regulated by a control valve.
STEAM SUPPLY
, FLOAT B THERMOSTATIC
GATE VALVE
(PLUG
Piping
Single
TYPE)
Coils
Fiqu~
122 illustrates a typical steam piping dia<
gram for coils used in either a high or medium
pressure system. If the return line is designed for
low pressure or vacuum conditions and for a pressure differential of 5 psi or greater from steam to
condensate return, a flash trap may be used.
Low pressure steam piping for a single coil is
illustrated in Fig. 123. This diagram shows an open
air relief located after the steam trap close to the
unit. This arrangement permits non-condensable
gases to vent to the atmosphere.
Fig1rt.e 12-f shows the f>iping layout for a steam
coil in a V;ICIIII~ system. h 15” check valve is used to
eq ilalile the vacuum across the steam trap.
CONDENSATE
GATE
,
DRIP
L I N E - , I
,
TO UNIT. REOUIRED ON
LOW PRESSURE GRAVITY
RETURN SYSTEMS.
VALVE-I
CONDENSATE
RETURN MAIN
Y
NOTES:
I. A bypass is necessary around trap and vaivcs
when
continuous operation is necessary.
2. Bypass to be the same size as trap orifice I)ut never
less than I/ inch.
1;1(;. 125 - DRIPI~ING
SIXAM SIJPPL\~
KI’TURN
1‘0
C O N D EN S A T E
I’ \K’l :I. I’I I’I NC; I)I:SI<;N
3-too
,-CONTROL VALVE
,-GATE
VALVE
(NOTE
2)
,-STRAINER
CONTROL
VALVE
.r- G A T E V A L V E
‘,
-REFER TO FIG. 125
WHEN DRIPPING STEAM SUPPLY
MAIN
TO
CONDENSATE
RETURN
CONDENSATE
RETURN
’
rIS” C H E C K V A L V E
FOR
NOTES:
1. Flange or union is located to facilitate coil removal.
2. A bypass is necessary around valves and strainer
when continuous operation is necessary.
3. Bypass to be the same size as valve port hut never
less than IA inch.
F1c.126 - BYI~ASSWITH
Dripping
Steam
Supply
MANUALCONTROL
“\~
BREAKING
+“PET COCK FOR
CONTINUOUS
IS’
DIRT LEG
VACUUM
VENT
CHECK
VALVE
(6”)
GATE
Main
FLOAT OR BUCKET
TRAP (NOTE 3 B 4)
A typical method of dripping the steam supply
main to the condensate return is shown in I;ig. 125.
/-Y--l
STEAM
SUPPLY
VALVE
CONDENSATE
NOTES:
1. Flange or union is located to facilitate coil removal.
2. When bypass control is required, see Fig. l-36.
3. Flash trap can be used if pressure differential between supply and condensate return exceeds 5 psi.
4. Coils with different presume drops require individual
traps.
5. Dirt pocket may be replaced by a strainer. If so, tee
on drop can be replaced by a reducing ell.
6. The petcock is not necessary with a bucket trap
or any trap which has provision for passing air. The
great majority of high pressure return mains terminate in hot wells or deaerators which vent the air.
FIG.
CHECK VALVE
/,-BUCKET
Steam
TRAP
c LOCATED BELOW
OUTLET,
DRAIN VALVE
(GATE.NOTE
2) ’
NOTES:
1. Flange or union is located
2. To prevent water hammer,
ting steam.
3. Do not exceed one foot of
and return main for each
ential.
FIG. 1!?-&NDENSA.I.E
to facilitate coil removal.
drain coil before admitlift between trap discharge
pound of pressure differ-
LIFT TOOVERHEAD
RETURN
128 -MULTIPLE
Bypass
COIL
HIGH
PRESSURE
PIPING
Control
Frequently a bypass with a manual control valve
is required on steam coils. The piping layout for a
control bypass with a plug type globe valve as the
manual control is shown in Fig. 126.
Lifting
Condensate
to
Return
Main
A typical layout for lifting condensate to an overhead return is described in I;ig. 1’7. The amount
of lift possible is determined by the pressure differential between the supply and return sides of the
system. The amount of lift is not to exceed one foot
for each pound of pressure
differential. The maximum lift sl~o~~ld not exceed 8 ft.
I
(:11,\1’~1‘1~~1<
I .
3-101
S’I‘I~.\XI I’II’lN(.
,/-CONTROL
(NOTES
N O T E I -._
.k/
.
\
CONTROL
VALVE
(NOTES283)
VALVE
283)
STRAlNER
8-1i’
,I
GATE
VALVE
STEAM
m-IS’CHECKV A L V E
WHEN
SUPPLY
SUPPLY
REFER TO FIG. 125
WHEN DRIPPING
SUPPLY TO RETURN.
DRIPPING
TO RETURN.
----THERMOSTATIC
,OPEN
TO
TRAP
THERMOSTATIC
(4”)
AIR RELIEF
ATMOSPHERE
TRAP
EQUALIZING
(4”)
VACUUM
CONDENSATE
&CHECK VALVE
TRAP (NOTE
TRAP
(NOTE
41
4)
GATE
VALVE
_i
r/
LCONDENSATE
RETURN
XOTES:
1. Flange or union is located to facilitate coil removal.
2. See Fig. 131 when control valve is omitted on multiple coils in parallel air flow.
3. When bypass control is required, see Fig. 126.
4. Coils with different pressure drops require individual
traps.
NOTES:
1. Flange or union is located to facilitate coil removal.
2. See Fig. 131 when control valve is omitted on multiple coils in parallel air flow.
3. When bypass control is required, see Fig. 126.
4. Coils with different pressure drops require individual
traps.
FIG.
130
-MULTIPLE
COIL
SYSTEMS
FIG .
129
-MULTIPLE
COIL
Low
PRESSURE
Low
PRESSURE
VhcuuM
P~privc
PIPING
FREEZE-UP
Piping Multiple Coils
Figures 117s tl21.u 131 show piping layouts for high
pressure, low pressure and vacuum systems with
multiple coils. If a control valve is not used, each
coil must have a separate steam trap as illustrated
in Fig. /31. This particular layout may be used for
a low pressure or vacuum system.
If the coils have different pressure drops or capacities, separate traps are required with or without a
control valve in the system.
Boiler Piping
I ; i g t l w Ii2 illustrates a suggested layout for a
steam plant. This diagram sho\vs parallel boilers
and a single boiler usirq a “Hartford Return Loop.”
PROTECTION
tVhen steam coils are used for tempering or preheating outdoor air, controls are required to prevent freezing of the coil.
I’n high, medium, low pressure and vacuum systems, an immersion thermostat is recommended to
protect the coil. This protection device controls the
fan motor and the outdoor air damper. The immersion thermostat is actuated when the steam supply
fails or when the condensate temperature drops
below a predetermined level, usually 120 F to 150 F.
‘I‘lie thermostat location is shown in Fig. 133.
The 15” check valve shown in the various pifiing
diagrams provides a means of equalizing the pressure within the coil when the steam supply shuts off.
This check valve is used in addition to the immersion thC‘rn1ostat.
The petcock for continuous venting removes non-condensable gases from the coil.
UNIT
1
/
1
I
STEAM SUPPLY
/’ MAIN
IMMERSION
THERMOSTAT
( SEE NOTE)
WHEN DRIPPING ST EAM SUPPLY
MAIN TO CONDENSATE RETURN
4
12”MIN.
b-d
CHECK VALVES
KOTE: Immersion thermostat is for control of outdoor air dampers and fan motor. Thermostat
closes damper and shuts ofi fan when condensate temperature drops below a predet rmined level.
”
NOTE:
Flange or
removal.
F1c;.131--Low
union
is
located
to
facilitate
coil
FIG. 133 -FREEZE-UP PROTECTIONFORHIGH,~IEDIUM,
Low PRESSURE,ANDVACUUM SYSTEMS
PREWJREOR Vxuuhf SYSTEM
STEAMTRAPS
below 165 F. Condensate drains are limited to a
5 ib pressure.
Non-condensable gases can restrict the How of condensate, causing coil freeze-up.
On a low pressure and vacuum steam heating
system, the immersion thermostat may be replaced
b y a condensate drain with a thermal element
(Fig. 134). The thermal element opens and drains
the coil when the condensate temperature drops
CONDENSATE
OUT
SCREEN
81 METAL
STEAM TRAP
SEE NOTE 8 DETAIL A
SITE DRAIN
NOTE:
Condensate
temperature
drain
drops
drains
below
coil when condensate
a predetermined level.
F1c.134 - FREEZE-UP PROTECTION FOR
AND VACUUM SYSTEMS
Low PRESSURE
‘l‘lie I’ollowing a r e g e n e r a l Icconimcntlations
ill
laying out systems hntlling outdoor air hclow 35 F:
I. Do n o t use ovcrluxtl returns lrom the hc;lting
unit.
2. I’rcssurc c o n t r o l s arc 11ot recomn~entletl si1ic.c
thy do liot n e c e s s a r i l y rellect actual contli-
tions. For example, it is pmsiblc for the coil to
Iwcome air huntl and have pressure hut no
steani. Also, the steam trap may be pluggd.
Prcssurc controls are slow acting by comprison with thermostatic controls.
3. USC ;L strainer in the supply line
ahcad of the trap.
;~ntl
2 dirt leg
Part 4
REFRIGERANTS, BRINES, OILS
4-l
CHAPTER 1. REFRIGERANTS
This chapter provides information concerning
the refrigeration cycles and characteristics of the
commonly used refrigerants and their selection for
use in air conditioning applications.
To provide .refrigeration, ,a refrigerant may be
utilized either:
1. In conjunction with a compressor, condenser
and evaporator in a compression cycle, or
2. With an absorbent in conjunction with an
absorber, generator, evaporator, and condenser
in an absorption cycle.
The refrigerant absorbs heat by evaporation generally at a low temperature and pressure level. Upon
condensing at a higher level, it rejects this heat to
an.1 available medium, usually water or air.
~_
a compression system the refrigerant vapor is
increased in pressure from evaporator to condenser
pressure by the use of a compressor.
In an absorption system the increase in pressure
is produced by heat supplied from steam or other
suitable hot fluid which circulates thru a coil of
pipe. The absorber-generator is analogous to a
compressor in that the absorber constitutes the
suction stroke and the generator the compression
stroke. The evaporator spray header corresponds
to the expansion valve. The evaporator and condenser are identical for both compression and absorption systems.
This chapter includes a discussion of the refrigeration cycle, refrigerant selection, and the commonly
used refrigerants as well as tables indicating their
characteristics and properties.
REFRIGERATION
ABSORPTION
CYCLES
CYCLE
The absorption refrigeration cycle utilizes two
phenomena:
1. The absorption solution (absorbent plus refrigerant) can absorb refrigerant vapor.
2. The refrigerant boils (flash cools itself) when
subjected to a lower pressure.
These two phenomena are used in the lithium
bromide absorption machine to obtain refrigeration by using the bromide as an absorbent and
water as a refrigerant.
Water is sprayed in an evaporator which is maintained at a high vacuum. A portion of the water
flashes and cools that which remains. The water
vapor is absorbed by a lithium bromide solution
in the absorber. The resulting solution is then
heated in the generator to drive off the water vapor
which is condensed in the condenser. The water
is returned to the evaporator, completing the cycle.
Figure 1 illustrates the absorption cycle. Figure 2
illustrates the cycle plotted on the equilibrium
diagram with numbered points representing pressures, temperatures, concentrations in the cycle.
On the lower left side of Fig. 1 is the absorber
partially filled with lithium bromide solution. On
the lower right side is the evaporator containing
water. A pipe connecting the shells is evacuated so
that no air is present. The lithium bromide begins
to absorb the water vapor; as the vapor is absorbed,
the water boils, generating more vapor and causing
the remainder of the water to be cooled.
LLEO
ER
I
SOLUTION
I
I
6
\
I
GENERATOR
PUMP
F IG . 1
- AB S O R P T I O N R E F R I G E R A T I O N
Since the water can vaporize more easily if it is
being sprayed, a pump is used to circulate the
water from the bottom of the evaporator to a
spray header at the top. An evaporator tube bundle
is located under the evaporator spray header; water
inside the tubes, returning from the air conditioning coils or other load, is Hash-cooled by the water
on the outside of the evaporator tubes. The lithium
bromide solution absorbs water vapor easier if it is
sprayed; therefore, a pump is used to circulate the
solution from the bottom of the absorber to a spray
header at the top of the absorber.
As the lithium bromide continues to absorb water
vapor, it becomes diluted, and its ability to absorb
additional water vapor decreases. The weak solution is pumped to the generator where heat is applied by steam or other suitable hot fluid in the
.
C YCLE
generator tube bundle to boil off the water vapor.
The solution is concentrated and returned to the
absorber. Since the weak solution going to the generator must be heated and the strong solution
coming from the generator must be cooled, a heat
exchanger is used in the solution circuit to conserve heat.
Water vapor boiled from the solution in the generator passes to the condenser to contact the relatively cold condenser tubes. The vapor condenses
in the condenser and returns to the evaporator so
that there is no loss of water in the cycle. Before
the condenser water goes thru the condenser tubes,
it passes thru a tube bundle located in the absorber.
Here it picks up the heat of dilution and the heat
of condensation which is generated as the solution
absorbs water vapor.
20.0
.
16.0
12.0
8.0
6.0
5.0
4.0 ,
3.5 .
3.c ,
2.: b
-
)
.
I
.
3
.
0
.
.4 0 .
.3 0
.
.2 5
.2 0
_I 8
.
.
,T
COMPRESSION
CYCLE
The compression refrigeration cycle utilizes two
phenomena:
1. The evaporation of a liquid rekigerant absorbs heat to lower the temperature ok its
surroundings.
2. The condensation of a refrigerant vapor rejects heat to raise the temperature of its
surroundings.
The cycle may be traced from any point in the
system. Figwe 3 is a schematic and Fig. f is a pressure-enthnlpy diagram of a compression cycle.
I’;\I<‘I‘
4-4
,I. I~E~III(;I:l~;\N’I‘S,
I1KINES,
O I L S
WATER-COOLED
LIQUID
STOP
FIG .
.~
3
REFRIGERANT
VALVE
- RECIPROCATING COMPRESSION REFRIGERATION CYCLE
Starting with the liquid refrigerant ahead of the
evaporator at point A in both Fig. 3 and 4, the
admission of liquid to the evaporator is controlled
by an automatic throttling device (expansion
valve) which is actuated by temperature and pressure. The refrigerant pressure is reduced across the
valve from condenser pressure, point A, to the evaporator pressure, point B. The valve acts as a boundary between the high and low sides of the system.
The pressure reduction allows the refrigerant to
boil or vaporize. To support boiling, heat from the
air or other medium to be cooled is transmitted to
the evaporator surface and into the boiling liquid
at a lower temperature. The refrigerant liquid and
vapor passing thru the evaporator coil continues to
absorb heat until it is completely evaporated, point
C. Superheating of the gas, controlled by the expansion valve, occurs from C to D.
The superheated gas is drawn thru the suction
line into the compressor cylinder. The downstroke
of the piston pulls a cylinder of gas thru the suction valve and compresses it on the upstroke, raising
its temperature and pressure to point E. The pressure produced causes the hot gas to flow to the condenser. The compressor discharge valve prevents
re-entry of compressed gas into the cylinder and
forms a boundry between the high and low sides.
In the condenser the condensing medium (air
or water) absorbs heat to condense the hot gas.
Liquid refrigerant is collected in receiver which may
be combined with or separate from condenser.
.
CONDENSING
EVAPORATION
ENTHALPY
FIG .
4
(ETUILB)
- P RESSURE-E NTHALPY DIAGRAM,
COMPRESSION
CYCLE
The liquid is then forced thru the liquid line to
the expansion valve A to repeat the cycle.
Liquid-Suction
Interchangers
Compressor ratings for Refrigerants 12 and 500
are generally based on 65 F actual suction gas’temperatures. When this suction gas temperature is
not obtained at the compressor, its rating must be
lowered by an appropriate multiplier. To develop
the full rating, the required superheat which is over
and above that available at the evaporator outlet
may be obtained by a liquid-suction interchanger.
I
ENTHALPY (BTU/LB)
SUPERHEATCF,
FIG. 5 - EFFECT
ON
LIQ~JID-SUCTION
INTERCHANGER
COMPRESSION CYCLE
OF
The effect of a liquid-suction interchanger on
the refrigeration cycle is shown on the pressureenthalpy diagram (Fig. 5).
The solid lines represent’ the basic cycle, while
the dashed lines represent the same cycle with a
liquid-suction interchanger. The useful refrigerating
effect with an interchanger is B’C rather than BC
as in the basic cycle.
Superheat increases the specific volume of the
suction vapor to reduce the total weight of refrigerant circulated for a given displacement. It
also increases the enthalpy of the vapor and may
improve compressor volumetric efficiency. Provided
thn heat absorbed represents useful refrigeration
L
I as liquid subcooling, the refrigerating effect
per pound of refrigerant circulated is increased.
With Refrigerants 22 and 717 volume increases
faster than refrigerating effect; hence, superheating
theoretically reduces capacity. With Refrigerants
12 and 500 the reverse is true, and superheating
theoretically reduces both the cfm per ton and the
power per ton. Figure 6 illustrates the loss due to
superheating of the refrigerant vapor and the gain
due to liquid subcooling. Net gain equals the gain
minus the loss.
REFRIGERANT PROPERTIES
Refrigerant characteristics have a bearing on system design, application and operation. A refrigerant is selected after an analysis of the required
characteristics and a matching of these require-
Courtesy of ASKE Data Uook
FIG. 6 - EFFECT
OF
LIQUID-SUCTION INTERCHANGER
CAPACITY
ON
ments with the specific properties of the available
refrigerants.
Significant refrigerant characteristics are:
1. Flammability and Toxicity as they pertain to
the safety of a refrigerant. The refrigerants
treated in this chapter are classified in
ASA H9.1 as Group 1, the least hazardous
relative to flammability and explosiveness.
The Underwriters Laboratories classification
with respect to toxicity puts these refrigerants
in Groups 4 to 6. The higher numbered groups
in this classification are the least toxic.
FigdYe 7 shows the structural formula of the
refrigerant compounds treated in this chapter.
The chlorine and fluorine elements in these
refrigerants make them the least hazardous
and least toxic respectively.
2. Miscibility of a refrigerant with compressor oil
aids in the return of oil from the evaporator
to the compressor crankcase in reciprocating
machine applications. Centrifugal units have
separate oil and refrigerant circuits.
Some refrigerants are highly miscible with
compressor oil. Refrigerants 12 and 500 and
lubricating oils are miscible in any proportion; Refrigerant 22 is less miscible. The effect
of miscibility in a refrigeration system is illustrated in Fig. 8. If Refrigerant I2 is placed
-
/
OIL ABSORBS ALL
OF THE REFRIGERANT
1.
CL
Refrigerant
11
CCl,,F
F
F
H
%+
0,L ABSORBS REFRIGERANT
EQUIVALENT
TO
67%
OF THE OIL WEIGHT
-
F
C
t
CL
OCL
Refrigerant 12
CC&F,
CL
Refrigerant 22
CHCIF.,
F
F
c
F
Refrigerant
c
‘c
41°
CL
CL
113
Refrigerant 114
C,,Cl,F,
- H - hydrogen
F - fluorine
%C13F3
C - carbon
Cl - c?lorine
F IG. 7 -STRUCTURAL
FIG.
F
‘c
+---ICL
CL
.
F
F
F
FORMULAS
FOR
REFRIGERANTS
in one vessel and lubricating oil in another
(Fig. Sn), and if both vessels are placed in a
common ambient temperature, all of the refrigerant migrates to the vessel containing oil
because of the absorption head of the oil.
Raising the oil temperature limits this migra.
tion. For instance, if the oil is at an ambient
temperature 20 degrees higher than the refrigerant or if the oil is heated by an immersion type heater to 20 degrees higher than the
refrigerant temperature (Fig. Sb), only 6’iy0
of the oil weight of the refrigerant becomes
dissolved in the oil.
3. Theoretical Horsepower Per 7’o~l of refrigeration for most refrigerants at air conditioning
temperature
(Table 1).
levels
is
R E F R I G E R A N T 12
OIL
b
approximately
the
same
4. Rate of Lectkngc of a refrigerant gas increases
directly with pressure and inversely with molecular weight. The prcssurc of a refrigerant
for a given saturated temperature increases
in the following order: Refrigerant 113, 11,
114, 12, 500, 22. The molecular weight of a
.
8
-MISCIBILITY
OF
REFRIGERANT
12
AND
OIL
refrigerant decreases in the following order:
Refrigerant 113, 114, 11, 12, 500, 22.
Molecular weight is directly related to vapor
specific volume; the higher the molecular
weight, the higher the specific volume.
5. Leak Detection of the refrigerant should be
simple and positive for purposes 01 maintenance, cost and safety. The use of a halide
torch makes it possible to detect and locate
minute leaks of the halogen refrigerants.
6. ~‘npol- Density influences the compressor dis-
placement and pipe sizing. High vapor density
accompanied by a reasonably high latent heat
of vaporization (low cfm per ton) is desirable
in a refrigerant. A low cfm per ton results in
compact equipment and smaller refrigerant
piping. Reciprocating refrigeration equipment
requires a relatively high vapor density refrigerant for optimum performance. Centrilugal compressors require a low vapor density
refrigerant for optimum efficiency at comparatively low tonnages. High vapor density
refrigerants are used with centrifugals of large
tonnage. The cfm per ton increases in the
following order: Refrigerant 22, 500, 12, 114,
11, 113 (Fig. 9).
Cost which is usually a consideration in all selections should not inlluence the choice of a refrig-
erant since it generally has little economic bearing
on the normal refrigeration system. Although Refrigerant 22 costs approximately twice as much as
Refrigerarlt I?, the compressor required is smaller,
tending to offset the additional cost of refrigerant.
39.5
CFM
n
1.98
CFM
2.68
CFM
3.14
CFM
non
500
I2
114
II
NOTE: Evaporator temp, 40 F
Condenser tcmp, 105 F
combined with the extended surl’ace rc:sults in ;I
significant increase in the over;lll heat transrcr rate.
When cooling coil is designed to permit 3 normal
pressure drop with Refrigerant 22, it may have a
rather large pressure drop with Refrigerant I2 or
500. In such a case the decreased performance for
Refrigerants 12 and 500 is due partially to the difference in their conclensing
film coefficients. In
addition, it is affectecl
by the pressure drop and
the resulting lower mean effective temperature
difference.
CONDENSERS
Frc.9-SUCTION VOLUMES OF REFRIGERANTS
(CFM/T~N)
l-6.
.T TRANSFER COMPARISONS
The value of the evaporating and condensing film
coefficient (Btu/sq ft/F) for Refrigerant 22 is greater
than that for Refrigerants 12 and 500. However, it
does not follow that cooling coils and condensers
can, therefore, be rated for higher capacities with
Refrigerant 22. The higher coefficient for Refrigerant 22 does not tell the complete story.
Various other factors should be considered:
1. Whether the heat exchanger is designed for
either Refrigerant 22 or Refrigerants 12 and 500.
2. Whether heat transfer is between refrigerant
and air or between refrigerant and water.
3. Whether tubes in a heat exchanger are prime
surface or extended surface (including the
amount of extended surface.)
The evaporating or condensing film coefficient is
0
f a number of factors which make up the total
ovei-all transfer rate for the heat exchanger. Other
factors involved are these:
1. Tube wall resistance (including extended surface, if any).
2. Air or water film coefficient.
3. Refrigerant pressure drop per circuit, which affects the mean effective temperature difference.
4. Surface ratio of tube outside surface to inside
surface.
5. Fouling factors (water-cooled condensers).
COOLING
COILS
Cooling coils using Refrigerant 22 normally provide a greater capacity than those using Refrigerant
12 or 500. A cooling coil, normally, has considerable
extended surface on the air side of the tubes. The
higher evaporating film coefficient of Refrigerant 2:!
Water-cooled condensers using Refrigerant 22
normally provide a greater capacity than those using
Refrigerant 12 or 500, depending on the fouling
factor used in its selection. There are a number of
reasons for this improvement other than the basically higher condensing film coefficient of the
refrigerant.
1. The water film coefficient is relatively high as
compared to air.
2. The extended surface or refrigerant side of
the exchanger tubes assures maximum transfer
rate and opt&urn balance between the inside
and outside surfaces.
With Refrigerant 12 or 500 the performance of
water-cooled condensers is otherwise not adversely
affected since the pressure drop in the shell is not
a consideration.
Air-cooled condensers using Refrigerant 22 normally provide a greater capacity than those using
Refrigerant 12 or 500.
Air-cooled condensers have considerable extended
surface on the air side of the tubes. The higher
evaporating film coefficient of Refrigerant 22, combined with the extended surface results in
significant increase in the overall heat transfer rate.
When an air-cooled condenser is designed to
allow for a normal pressure drop with Refrigerant
22, it may have a considerable pressure drop when.
used with Refrigerant 500 or 12. In such a case the
decreased performance for Refrigerants 500 and 12
is not due entirely to the difference in their condensing film coefficients, but is affected also by pressure drop and the resulting lower mean effective
temperature difference.
Normally, evaporative condensers use prime surface tubing (without extended surface on the outside of tubes). Where the unit is designed for
Refrigerant 12 or 500, there is no significant increase in capacity when using Refrigerant 22.
If the unit is tlcsigried for Kefrigerant 22 (smaller
tubing or longer circuits) , and is used with Refrigerant 12 or 500, the pressure drop is sufficient to
reduce the rating. SLI& ;I chip m;iy create the
impression that the increased rating for Refrigerant
22 is due to the condensing film, whereas actually
the performance for Refrigerant 12 or 500 rlccreases
due to coil design.
REFRIGERANT
COMPRESSION
ABSORPTION
CYCLE
Water 2s a refrigerant and lithium bromide as
;in absorbent are utilized in the basic absorption
refrigeration cycle. The refrigerant slw~~ltl possess
the same tlcsirablc qualities as those for a compression system. In addition, it should be suitable for
use with an absorbent, so selected that:
SELECTION
CYCLE
The choice of a refrigerant for a compression
system is limited by:
1. Economics,
2. Equipment type and size
3. Application
The manufacturer of the refrigeration compressor
generally preselects the refrigerant to result in optimum owning cost. The specific refrigerant is determined by the type and size of the equipment.
To minimize the number of reciprocating compressor sizes required to fill out a line, the manufacturer rates each size for several refrigerants of
relatively dense vapor such as Refrigerants 12, 500
and 22. In effect, this increases the number of units
offered without adding sizes.
Centrifugal compressors at comparatively low
tonnages require a high vapor volume refrigerant
such as Refrigerant 113 or 11 to maintain optimum
efficiency. For most sizes Refrigerant 114 or 12 can
be used to obtain greater capacities. Refrigerants
500 and 22 are used with specially built centriEugals to obtain the highest capacities.
The refrigerant selected depends on the type of
application. Air-cooled condensers may not use certain refrigerants because of the design condensing
temperature required and corresponding limitations
on compressor head pressure.
The temperature-pressure relationship of a refrigerant is of considerable importance in low temperature applications. If the evaporator pressure is
comparatively low for the required evaporator temperature, the volume of vapor to be handled by the
compressor is excessive. If the evaporator pressure
is comparatively high for the required evaporator
temperature, the system pressures are high.
The refrigerants that have been mentioned are
the halogens (fluorinated hydrocarbon compounds),
except for Refrigerant 500 which is an azeotropic
mixture of two Auorinated hydrocarbons. The mixture does not separate into its component refrigerants with a change of temperature or pressure. It
has its own fixed thermodynamic properties which
are unlike either of its components.
.
30
IN.
4ES.
3
0
NOTE:
-7
Absorption machine at full load, tising lithilim
bromide as absorbent.
FIG. 10
- P RESSURES
AND
T EMPERATURES
OF A
T Y P I C A L ABSORPTION M A C H I N E
1. The vapor pressures of the refrigerant and absorbent at the generator are different.
2. The temperature-pressure relations are consistent with practical absorber and generator
temperatures and pressures. Figure 10 shows
absolute pressures and temperatures existing in
a typical absorption machine at full load.
3. The refrigerant has a high solubility in the
absorbent at absorber temperature and pressure and a low solubility at generator temperature and pressure.
4. The refrigerant and absorbent together are
stable within the evaporator-generator range
of temperatures. Normally, the absorbent must
remain liquid at absorber and generator temperatures and pressures. It should have a low
specific heat, surface tension and viscosity and
must be neutral to the materials used in the
equipment.
ant1
tlicir c.ll;l~;icteristi(,s.
‘l’ir/~/f,.r I” l o
7
l i s t
tllc
temperatures.
TABLE l-COMPARATIVE DATA OF REFUGERANTS
REFRIGERANT NUMBER
(AR1
DESIGNATION)
11
Trichloromonofluoromethane
Chemical Name
Chemical Formula
Molecular wt
Gas Constant, R !ft-lb/lb-R)
Boiling Point at 1 atm (F)
Dichlorodifluoromethane
CCl3F
CC12F2
I3 7 . 3 8
120.93
11.25
388.0
635.0
Specific Heat of liquid, 86 F
.220
.235
Specific Heat of Vapor, Cp
6 0 F at 1 atm
*
.146
-.
ific Heat of Vapor, Cv
60 F at 1 atm
Triehlorotrifluoroethane
CHClFz
CC12F-CCIFz
86.48
187.39
.156
.171
.127
*
.145
.151
1.12
Ratio of Specific Heats
liquid, 105 F
2.04
1.55
2.14
1.47
.
105 F
Net Refrigerating Effect (Btu/lb)
4 0 F-105 F fno subcoolina)
Cycle Efficiency (yo Carnot Cycle)
40 F-105 F
Solubility
of Water in Refrigerant
Miscibility with Oil
Toxic Concentration (yo by vol)
Odor
Warning Properties
Explosive Range (yo by vol)
c
‘y G r o u p , U.I..
\I Group, ASA B9.1
To. ‘c Decomposition Products
Viscositv
(centiooises)
Sot&ted liquid
95 F
I
2.04
50 F
!
liquid Circulated,
(Ib/min/ton)
4( 1 F-105 F
Theore tical Displacement, 40 F-105 F
_,_I
.
(cu ftlminltont
0.52
2.55
7.03
25.7
7.12
23.05
51.67
141.25
11.74
38.79
03.72
227.65
0.84
2.66
11.58
67.56
49.13
66.44
54.54
90.5
03.2
81.8
87.5
Negligible
1.59
1.77
Negligible
1.65
1
2.10
*
Miscible
Above 1 0 %
Negligible
Miscible
Above 2070
Negligible
Ethereal,
odorless w h e n
m i x e dw i t ha i r
Same asR 11
Same or R 11
SameasRll
Same as R 11
SameasRll
None
NOW
5
Non-2
NOW
NOW
None
1
6
1
None
NOW
NOW
NOW
6
NOW3
NOW
5A
Yes
Ye*
Ye*
Yes
Miscible
*
Limited
*
5A
1
4-5
1
I
.0121
.0105
I
,
1
Ye*
1
Ye*
.3420
.3272
.0108
.0109
.OlOO
.Olll
.2150
.2100
*
*
*
.0124
I
*
*
*
*
*
I
.0046
.0051
.0063
.0040
.0435
. 0 4 21
.0056
.0057
.0059
2.96
4.07
3.02
3.66
4.62
3.35
3.14
1.98
9.16
2.69
0.736
0.75
0.70
0.722
0.747
6.39
6.29
6.74
6.52
16.1
Theoretical Horsepower Per Ton
40 F-105 F
0.676
Coefficient of Performance
40 F-105 F (4.7l/hp
ton)
per
6.95
Cost Compared With R 11
1.00
*Datanotavailableornotapplicable.
1.13
+
1.51
Vapor at 1 atm
I
1.09
I
1.84
I
50 F
T h e r m a ,C o n d . . c l : ~ : . ~I L ,
-254
.218
1.18
OF
40 F
170.93
9.04
30.4
-137
*
1.14
1.61
73.a% CClzF2
26.27~ CHKHFz
99.29
15.57
-28.0
.149
1.11
Saturation Pressure (psia)
at: - 5 0 F
Aseotrope of
Dichlorodifluoromethane
and
Difluoroethane
cZc12F4
117.6
-31
417.4
495.0
Ratio 4 =K (86 F at 1 atm)
Vapor, Cp, 40 F sot. press.
liquid Head (ft), 1 psi at 105 F
50/o
0 ltchlorotetraf luoroethane
8.25
-41.4
-256
204.8
716.0
.335
.130
*
114
113
17.87
-21.62
-252
233.6
597.0
-I68
22
Monochlorodifluoromethone
12.70
74.7
Freerina Point at 1 otm (F)
Critical
Temperature 6)
C r i t i c a l PreSSWe
bid
/
12
.
39.5
6.3 1
I
1.57
2.77
2.15
2.97
2.00
4-10
-
-
-
-
-
:1/v\
v
]'\I< I‘ 1.
_-. - -~. ..--
--__-
J~J~:J~1~1O1~lI.\NJ'.S,
.
JIl~INI:.S.
OILS
TABLE 2-PROPERTIES OF REFRIGERANT 12, LIQUID AND SATURATED VAPOR
TEMP
(F)
PRESSURE
W/r
bmtute
P
-100
-9
- 9
- 9
- 9
- 9
-
8
6
4
2
0
aa
- 06
- a 4
- 82
-
80
78
76
74
72
70
68
66
64
62
6 0
58
-
56
5 4
5 2
5 0
-
48
- 46
- 44
- 42
- 40
- 38
- 36
- 34
- 32
- 30
- 28
- 26
- 24
- 2 2
- 2 0
- 1.3
-
16
14
12
- 1 0
1.4280
1.5381
1.6551
1.7794
I.9112
2.0509
2.1988
2.3554
2.5210
2.6960
2.8807
3.0756
3 . 2 8 1I
3.4975
3.7254
3.9651
4 . 2 172
4.4819
4.7599
5.0516
5.3575
5.6780
6.0137
6.3650
6.7326
7.1168
7 . 5 1 a3
7.9375
a.3751
a.8316
9.3076
9.8035
10.320
10.858
11.417
11.999
12.604
13.233
13.886
14.564
15.267
15.996
16.753
17.536
18.348
1 9 . 1 a9
20.059
20.960
21.891
22.854
23.849
24.878
25.939
27.036
28.167
29.335
30.539
31.780
33.060
34.378
35.736
37.135
38.574
40.056
41.580
n.)
Gage
P
7.0138*
'6.7896'
'6.5514*
#6.2984'
'6.0301'
15.7456'
!5.4443*
!5.1255*
!4.7884*
!4.4321*
!4.0560*
!3.6592+
!3.2409*
z2.8002*
t2.3362'
!1.8482*
11.3350*
20.7959*
20.2299'
19.6360*
19.0133*
18.3607'
17.67731
16.9619*
16.2136*
15.4313*
14.6139*
13.7603*
12.8693'
11.9399'
10.9709*
9 . 9 6 1 1'
8.909*
7.814*
6.675'
5.4901
4.259*
2.979'
1.649'
0.270+
0.57I
1.300
2.057
2.840
3.652
4.493
5.363
6.264
7.195
8.158
9.153
lo.182
11.243
12.340
13.471
14.639
15.843
17.084
18.364
19.682
21.040
22.439
23.878
25.360
26.884
rc
VOLUME
(cu I
Liquid
vt
3)
lC2p.X
l-
vg
0 . 0 0 9 9 8 5 2;!.I 64
. 0 1 0 0 0 2 2(I.682
. 0 1 0 0 2 0 1s I.316
. 0 1 0 0 3 7 lf I.057
. o1 0 0 5 5
Id i.895
0 ~.010073
1: i.821
.010091
111.828
.010109
1.1.908
.oio128
1:I.056
.010146
1:2.226
C I.010164
111.533
.010183
I( 1.852
. 0 1 0 2 0 2 l( I.218
.010221
?.6290
.010240
?.oao2
C LO10259
3.5687
.01027a
3.0916
.01029a
7.6462
.010317
7.2302
.010337
6.8412
(1.010357
6.4774
.010377
6.1367
.010397
5.8176
.010417
5.5184
.010438
5.2377
( I.010459
4.9742
.010479
4.7267
.010500
4.4940
.010521
4.2751
.010543
4.0691
,3.010564
3.8750
.010586
3.6922
.010607
3.5198
.010629
3.3571
.010651
3.2035
0.010674
3.0585
.010696
2.9214
.010719
2.7917
.010741
2.6691
.010764
2.5529
0.010788
2.4429
.oioali
2.3387
.010834
2.2399
.010858
2.1461
.010882
2.0572
0.010906
1.9727
.010931
1.8924
.010955
1.8161
.oio9ac
1.7436
.011005
1.6745
0.011030
1.6089
.011056
1.5463
.oiloa2
1.4867
.011107
1.4299
.011134
1.3758
0.011 I b C
1.3241
.011187
1.2748
.011214
I.2278
.Oll241
I.1828
.0112bE
1.1399
0.0 1 I 29C
I.0988
.01132A
1.0596
.01135;
1.0220
.01138(
0.9861
.01140(
0.9517
'Inches of mercury below one atmosphere.
DENSITY
fIb/cu ft)
T
Liquid
l/Vf
Vapor
119
LO45119
.048352
.051769
.055379
.059ia9
I.063207
.067441
.071900
.076591
.081525
93.690
93.493
93.296
93.098
92.899
92.699
92.499
92.298
92.096
91.893
91.689
91.485
91.280
91.074
90.867
90.659
90.450
90.240
90.030
89.818
89.606
89.392
89.178
0.32696
.34231
.35820
.37466
.39171
0.40934
.4275a
.44645
.46595
.48611
0.50693
.52843
.55063
.57354
.59718
0.62156
.64670
.67263
.69934
.72687
0.75523
.78443
.81449
.84544
.87729
88.746
88.529
88.310
88.091
87.870
87.649
Liquid
hf
- 12.466
- 12.055
-1 1 . 6 4 4
-11.233
-10.821
- 10.409
- 9.9971
- 9.5845
- 9.1717
- a.7586
- a.3451
).086708
.092151
- 7.9314
- 7.5173
.097863
- 7.1029
.I0385
.I1013
- 6.6881
).11670
- 6.2730
- 5.8574
.12359
.I3078
- 5.4416
.13831
- 5.0254
- 4.6088
.14617
- 4.1919
3.15438
- 3.7745
.16295
.17189
- 3.3567
- 2.9386
.I8121
- 2.5200
.19092
- 2.1011
0.20104
.21157 - 1.6817
.22252
- 1.2619
- 0.8417
.23391
- 0.4211
.24576
0.0000
0.25806
.27084
0.4215
.28411
0.8434
1.2659
.29788
.312lb
I .baa7
00.15
99.978
99.803
99.627
99.451
99.274
99.097
98.919
98.740
98.561
98.382
98.201
98.021
97.839
97.657
97.475
97.292
97.108
96.924
96.739
96.553
96.367
96.180
95.993
95.804
95.616
95.426
95.236
95.045
94.854
94.661
94.469
94.275
94.081
93.886
88.962
t
0.91006
.94377
.97843
1.0141
1.0507
T-
ENTHALPY
(B stu/lb)
I
2.1120
2.5358
2.9601
3.3848
3.8100
4.2357
4.661a
5.0885
5.5157
5.9434
6.3716
6.8003
7.2296
7.6594
8.0898
8.5207
1 8.9521
9.3843
T-i atent
‘+a
7 ‘a.71 4
7 '8.524
7 '8.334
7 '8.144
7 '7.954
7 '7.764
7'7.574
7 '7.384
7 '7.194
7 '7.003
7 '6.812
7'6.620
7'6.429
i '6.238
7'6.046
7 '5.853
7'5.660
i' 5 . 4 6 7
;' 5 . 2 7 3
, '5.080
i'4.885
j7 4 . 6 9 1
74.495
74.299
74.103
73.906
73.709
73.511
73.312
7 3 . 1 12
72.913
72.712
72.511
72.309
72.106
71.903
71.698
71.494
71.288
vapor
h,
66.248
66.469
66.690
66.911
67.133
67.355
67.577
67.799
68.022
68.244
68.467
68.689
68.912
69.135
69.358
69.580
69.803
70.025
70.248
70.471
70.693
70.916
71.138
71.360
71.583
70.874
70.666
70.456
70.246
70.036
69.824
69.611
69.397
69.183
68.967
68.750
68.533
68.314
68.094
67.873
71.805
72.027
72.249
72.470
72.691
72.913
73.134
73.354
73.575
73.795
74.015
74.234
74.454
74.673
74.891
75.110
75.328
75.545
75.762
75.979
76.196
76.411
76.627
76.842
77.057
77.271
77.485
77.698
77.911
78.123
67.651
67.428
67.203
66.977
66.750
66.522
66.293
66.061
65.829
65.596
78.335
78.546
78.757
78.966
79.176
79.385
79.593
79.800
80.007
80.214
71.081
ENTROPY
(Btu/l
R)
Liquid
If
-0.032005
- .030866
- .029733
- .028606
- .027484
-0.026367
- .025256
- .024150
- .023049
- .021953
Vapor
Ig
1.1 a683
.18623
.18565
.18508
‘18452
I.18398
.I8345
.I8293
.18242
.18192
1.18143
-0.020862
- .019776
.18096
- .018695
.l8050
- .017619
.I8004
- .016547
.I7960
-0.015481
I.17916
- .014418
.17874
- .013361
.17833
- .oi23oa
.17792
- .011259
.17753
-0.010214 3.17714
- .009174
.I7676
- .008139
.I7639
- .007107
.17603
.17568
- .006080
-0.005056 0.17533
.17500
- .004037
- .003022
.17467
.17435
- .002011
- .001003
.17403
0.17323
0.000000
.001000
.17343
.001995
.17313
.I7285
.002988
.003976
.I7257
0.004961
0.17229
.005942
.17203
.006919
.17177
.007894
.17151
.ooaa64
.17126
0.009831
0.17102
.010795
.17078
.17055
.011755
.012712
.17032
.013666
.I7010
0.014617
0.16989
.015564
.16967
.016508
.16947
.017445
.16927
.01838E
.16907
0.019323
0.16888
.0202x
.16869
.021184
*lb851
.02211c
-16833
.02303:
.I6815
0 . 0 2 3 9 5 4 0.16798
.024871
.I6782
.I6765
.02578(
.02669(
.16750
.0276Ot
.16734
0.02851:
0.16719
.I6704
.02942C
.03032:
.I6690
.I6676
.031221
.0321lt
.I6662
.EMP
(F)
t
--1oo
-
96
96
94
92
90
88
86
04
82
80
78
76
74
72
70
66
66 ,
64
62
60
5 8
56
54
5 2
50
48
46
44
42
-
40
38
36
34
32
-
30
2 8
26
24
22
20
1 8
16
-
14
-
1 2
-
10
8
-
6
4
2
0
2
4
6
a
10
12
14
lb
18
20
22
24
26
20
(iI I.\ I’ I I,.I< I.
1<1,:L’i<
4-11
(;I-I<.\x I’S
TABLE Z-PROPERTIES OF REFRlGERANT
TEMP
(F)
PRESSURE
in.)
(lb/
ibrolure
P
I
r
Gage
P
28.452
30.064
31.721
33.424
35.174
VOLUME
(cu f
3)
-I-
12, LIQUID
v(
o.oii43a
.0114ba
.011497
.011527
.011557
vg
3 . 9 1a 8 0
.a8725
.a5702
.a2803
.a0023
DENSITY
( l b / c' U f f)
Liquid
Vapor
1 lvt
1 i-d
37.426
1.0884
37.202
1.1271
36.977
1.1668
36.751
1.2077
16.524
1.2496
Liquid
vapor
-r
AND SATURATED VAPOR (Contd)
I--
ENTHALPY
Btu/lbJ
ENTROPY
Ib- RI
(BfU/
‘EMP
T-7 (F)
Liquid
hr
15.058
15.500
15.942
lb.384
lb.828
hfs
65.361
65.124
64.886
64.647
64.406
Vapor
he
80.419
80.624
80.828
81.031
81.234
Liquid
‘if
0.033013
.033905
.034796
.035683
.036569
Pg
0.16648
.16635
.16622
.lbblO
.16598
Latent
Vapor
I
30
32
34
36
38
43.148
44.760
46.4 17
48.120
49.870
40
42
44
46
48
51.667
53.513
55.407
57.352
59.347
36.971
38.817
40.711
42.656
44.65 1
0.011588
.011619
.011650
.oi 1682
.011714
3.77357
.7479a
.72341
.69982
.67715
36.296
86.066
35.836
35.604
35.371
1.2927
1.3369
1.3823
1.4289
1.4768
17.273
17.718
18.164
18.611
19.059
64.163
63.919
63.673
63.426
63.177
81.436
81.637
81.837
82.037
82.236
0.037453
.038334
.039213
.040091
.040966
0.16586
.I6574
.16562
.16551
.16540
40
42
44
46
48
50
52
54
56
58
61.394
63.494
65.646
67.853
70.115
46.698
48.798
50.950
53.157
55.419
0.011746
.011779
.oiiali
.oiia45
.oiia79
0.65537
.63444
.61431
.59495
.57632
35.136
34.900
34.663
34.425
34.1 a5
1.5258
1.5762
1.6278
I .6808
1.7352
19.507
19.957
20.408
20.859
21.312
62.926
62.673
62.418
62.162
61.903
82.433
82.630
82.826
83.021
83.215
0.041839
.042711
.0435ai
.044449
.045316
0.16530
.16519
.I6509
.I6499
.16489
50
52
54
56
58
60
62
64
66
72.433
74.807
77.239
79.729
82.279
57.737
60.111
62.543
65.033
67.583
0.011913
.011947
0.11982
.012017
.012053
0.55839
.54112
.52450
.5084a
.49305
83.944
83.701
83.457
83.212
82.965
1.7909
1.8480
1.9066
1.9666
2.0282
21.766
22.221
22.676
23.133
23.591
61.643
61.380
61.116
60.849
60.580
83.409
83.601
83.792
83.982
84.171
0.046180
.047044
.047905
.048765
.049624
0.16479
.I6470
.16460
.16451
.I6442
60
62
64
66
68
.O
72
74
76
78
84.888
87.559
90.292
93.087
95.946
70.192
72.863
75.596
78.391
81.250
0.012089
.012116
.012163
.012201
.012239
0.4781 a
.46383
.45000
.43bbb
.4237a
82.717
82.467
82.215
81.962
81.707
2.0913
2.1559
2.2222
2.2901
2.3597
24.050
24.511
24.973
25.435
25.899
60.309
60.035
59.759
59.481
59.20 1
84.359
84.546
84.732
84.916
85.100
0.050482
.051338
.052193
.053047
.053900
0.16434
.I6425
.16417
.16408
.16400
70
72
74
76
78
80
a2
a4
a6
aa
98.870
101.86
104.92
108.04
111.23
84.174
87.16
90.22
93.34
96.53
0.012277
.012316
.012356
0.12396
0.12437
0 . 4 11 3 5
.39935
.3a776
.37657
.36575
81.450
al.192
80.932
80.671
80.407
2.4310
2.504 I
2.5789
2.6556
2.7341
26.365
26.832
27.300
27.769
28.241
58.917
58.631
58.343
58.052
57.757
85.282
85.463
85.643
85.821
85.998
0.054751
.055602
.056452
.057301
. 0 5 81 4 9
0.16392
.I6384
.lb376
.I6368
.16360
a0
a2
a4
86
aa
90
92
94
96
98
114.49
117.82
121.22
124.70
128.24
99.79
103.12
106.52
110.00
113.54
0.012178
.012520
.Ol2562
.012605
.012649
0.35529
.345ia
.33540
.32594
. 3l b 7 9
80.142
79.874
79.605
79.334
79.061
2.8146
2.8970
2.9815
3.0680
3.1566
28.713
29.187
29,663
30.140
30.619
57.461
57.161
56.858
56.551
56.242
86.174
86.348
86.521
86.691
86.861
0.058997
.059844
.060690
.061536
.062381
0.16353
.16345
.I6338
.16330
.I6323
90
92
94
96
98
1oc
102
104
1OC
106
131.86
135.56
139.33
143.18
147.11
117.16
120.86
124.63
128.48
132.41
0.012693
.012738
.012783
.012829
.012876
0.30794
.29937
.29106
.28303
.27524
78.785
78.508
78.228
77.946
77.662
3.2474
3.3404
3.4357
3.5333
3.6332
31.100
31.583
32.067
32.553
33.041
55.929
55.613
55.293
54.970
54.643
87.029
87.196
87.360
87.523
87.684
0.063227
.064072
.064916
.065761
.066606
0.16315
.16308
.16301
.16293
.I6286
100
102
104
106
108
11c
111
114
11c
118
151.11
155.19
159.36
163.61
167.94
136.41
140.49
144.66
148.91
153.24
0.012924
.012972
.013022
. O 13072
.013123
0.26769
.26037
.2532a
.24641
.23974
77.376
77.087
76.795
76.501
76.205
3.7357
3.8406
3.9482
4.0584
4.1713
33.531
34.023
34.5 17
35.014
35.512
54.313
53.978
53.639
53.296
52.949
87.844
88.001
88.156
88.310
88.461
0.067451
.068296
.069141
.069987
.070833
0.16279
.16271
.16264
.I6256
.16249
110
112
114
116
118
12c
4
126
128
172.35
176.85
181.43
186.10
190.86
157.65
162.15
166.73
171.40
176.16
0.013174
.013227
.oi32ao
.013335
.013390
0.23326
.22698
.22089
.21497
.20922
75.906
75.604
75.299
74,991
74.680
4.2870
4.4056
4.5272
4.6518
4.7796
36.013
36.516
37.02 1
37.529
38.040
52.597
52.241
51.881
51.515
51.144
88.610
88.757
88.902
89.044
89.184
0.071680
.072528
.073376
.074225
.075075
0.16241
.16234
.I6226
.I6218
.1621C
120
122
124
126
128
13c
13:
134
13t
13s
195.71
200.64
205.67
2 10.79
216.01
181.01
185.94
190.97
196.09
201.31
0.013447
.013504
.013563
.013623
.013684
0.20364
.I9821
.19294
.I8782
.I8283
74.367
74.050
73.729
73.406
73.079
4.9107
5.0451
5.1829
5.3244
5.4695
38.553
39.069
39.588
40.110
40.634
50.768
50.387
50.000
49.608
49.210
89.321
89.456
89.588
89.718
89.844
0.075927
.076779
.077623
.078489
.079346
0.16202
.16194
.i618
.I6177
.1616E
130
132
134
136
138
14t
14:
141
141
14z
221.32
226.72
232.22
237.82
243.51
206.62
212.02
217.52
223.12
228.81
0.013746
.oi3aio
.oi3874
.013941
.014008
0.17799
.17327
.I6868
.16422
.I5987
72.748
72.413
72.075
71.732
71.386
5.6184
5.7713
5.9283
6.0895
6.2551
41.162
41.693
42.227
42.765
43.306
48.805
48.394
47.977
47.553
47.122
89.967
90.087
90.204
90.31 a
90.428
0.080205
.081065
.oai92a
.082794
.083661
0.16155
.1615C
.1614C
.1613C
.1612C
140
142
144
146
148
1%
15:
15‘
15(
151
249.3 1
255.20
261.20
267.30
273.5 1
234.61
240.50
246.50
252.60
258.81
0.014078
.014148
.014221
.014295
.014371
0.15564
.15151
.I4750
.I4358
.I3976
71.035
70.679
70.319
69.954
69.584
6.4252
6 . 6 0 01
6.7799
6.9648
7.1551
43.850
44.399
44.951
45.508
46.068
46.684
46.238
45.784
45.322
44.852
90.534
90.637
90.735
90.830
90.920
0.084531
.085404
.086280
.087159
.088041
0.16llC
.1609(
.1608t
.1607i
b o .bl!
150
152
154
156
15a
16(
279.82
265.12
0.014449
0.13604
69.209
7.3509
46.633
44.373
9 1.006
0.088927
0.1605:
160
-
-
30
32
34
36
38
TABLE
TEMf
(F)
t
-40
- 3 8
- 3 6
- 3 4
- 3 2
- 3 0
- 2 8
- 2 6
- 2 4
- 2 2
- 2 0
- 1 8
-16
- 1 4
- 12
- 1 0
8
6
- 4
2
0
2
4
6
8
10
12
14
lb
18
20
2 2
24
26
28
30
32
34
36
38
40
42
44
46
\48
50
52
54
56
58
60
61
64
66
68
7a
71
74
76
78
BC
81
84
8t
a8
S-PROPERTIES OF REFRIGERANT 500, LIQUID AND SATURATED VAPOR
PRESSURE
(Ib/rq in.)
VOL iGdE
(cv 1 k/l b)
.iquid
Vf
15.45
16.21
17.01
17.84
18.70
19.59
20.5 I
21.47
22.46
23.49
24.55
25.65
26.79
27.96
29.18
30.43
31.73
33.06
34.45
35.88
37.35
38.86
40.42
42.03
43.69
45.40
47.15
48.96
50.82
52.74
54.71
56.72
58.80
0.75
1.51
2.3 1
3.14
4.00
4.89
5.81
6.77
7.76
a.79
9.85
10.95
12.09
13.26
14.48
15.73
17.03
18.36
19.75
21.18
22.65
24.16
25.72
27.33
28.99
30.70
32.45
34.26
36.12
38.04
40.0 1
42.02
44.10
-/
67.71 I 53.01
70.09
15 5 . 3 9
72.52 1 57.82
75.02 1 60.32
77.57
62.87
80.22
65.52
I . 0 11 9
.OllP
.0119
.0119
.0120
I.01 20
.0120
.0120
.0121
.0121
P--
vapor
vg
Liquid
l/vt
4.0757
3.8829
3.6992
3.5276
3.3671
3.2121
3.0674
2.9302
2.8020
2.6788
84.37
84.19
84.00
83.82
83.63
I.0121
.0121
.0122
.0122
.0122
I.0123
.0123
.0123
.0123
.0124
3.0124
.0124
.0125
.0125
.0125
0.0126
.0126
.0126
.0127
.0127
0.0127
.0128
.0128
.0128
.0129
2.5622
2.4520
2.3477
2.2491
2.1548
2.0657
1.9807
1.9004
1.8238
1.7507
1.6818
1.6155
1.5530
1.4929
1.4362
1.3813
1.3292
1.2796
1.2325
1.1873
1.1440
1.1027
1.0631
1.0254
0.9892
82.51
82.32
82.13
81.94
81.75
0.0129
.0129
.0130
.0130
.0130
0.0131
.0131
.0131
.0132
.0132
0.0133
.0133
.0133
.0134
.0134
0.0135
.0135
.0135
.0136
.0136
0.0137
.0137
.0138
.0138
.0139
0.0139
.0139
.0140
.0140
.0141
0.9545
.9212
.8894
.8591
.a298
0.8017
.7747
.7490
.7241
.7002
0.6774
.6554
.6343
.6138
.5943
0.5756
.5575
s400
.5231
.5069
0.4914
.4763
.4618
.4479
.4344
0.4213
.4087
.3967
.3849
.3735
*Inches of mercury below one atmosphere.
83.45
83.26
83.07
82.89
82.70
81.56
81.37
81.17
80.98
80.78
80.59
80.39
80.20
80.00
79.80
79.60
79.40
79.20
79.00
78.80
78.59
78.39
78.18
77.98
77.77
77.56
77.35
77.14
76.93
76.72
76.50
76.29
76.07
75.86
75.64
75.42
75.20
74.97
74.75
74.52
74.30
74.07
73.04
73.61
73.38
73.14
72.91
72.67
72.43
72.19
71.95
71.70
71.46
71.21
70.96
IY
ff)
vapor
l/Q
0.2454
.2575
.2703
.2835
.2970
0.3113
.3260
.3413
.3569
.3733
0.3903
.4078
.4260
.4446
.4641
0.4841
.5049
.5262
.5483
.5712
0.5946
.6190
.6439
.6698
.6963
0.7239
.7523
.7815
.a114
.8422
0.8741
0.9069
0.9407
0.9752
1.0109
1.0476
1.0855
1.1244
1.1640
1.2051
1.2473
1.2908
1.3352
1.3811
1.4282
1.4763
1.5258
1.5764
1.6291
1.6826
1.7374
1.7939
1.8518
1.9116
1.9727
2.0350
2.0996
2.1655
2.2328
2.3021
2.3735
2.4469
2.5210
2.5981
2.6772
r
.RO
ENTHALPY
Mu/lb)
liquid
hf
0.00
0 . 51
1.02
1.54
2.05
2.57
3.09
3.60
4.13
4.65
5.17
5.70
6.22
6.75
7.27
7.80
a.33
8.87
9.40
9.94
10.47
11.01
11.55
12.09
12.63
13.17
13.71
14.26
14.81
15.35
15.91
16.45
17.01
17.56
18.12
18.67
19.22
19.79
20.34
20.91
21.47
22.03
22.60
23.16
23.75
24.31
24.88
25.46
26.04
26.59
27.19
27.76
28.33
28.92
29.52
30.10
30.68
31.27
31.87
32.47
33.06
33.63
34.24
34.84
35.44
EN1
PY
(mu /lb.-R)
Latent
Vapor
b
b
89.91
89.68
89.45
89.21
88.97
88.73
88.49
88.25
88.00
87.75
87.50
87.25
87.00
86.74
84.49
06.23
85.97
85.70
85.44
85.17
84.90
84.63
84.35
84.08
83.80
83.52
83.24
82.95
02.66
82.37
82.07
81.78
81.48
81.18
80.87
80.57
80.26
79.94
79.63
79.31
78.99
78.67
78.34
78.01
77.66
77.33
76.99
76.64
76.29
75.96
75.59
75.24
74.89
74.51
74.13
73.76
73.39
73.01
72.62
72.22
71.83
71.46
71.04
70.63
70.22
89.91
90.19
90.47
90.75
91.02
91.30
91.58
91.85
92.13
92.40
92.67
92.95
93.22
93.49
93.76
Liquid
St
0.00000
94.03
94.30
94.57
94.84
95.11
.00122
.00243
.00365
.00485
0.00606
.00725
.00845
.00963
.01083
0.01202
.01322
.01438
.01557
.01674
0.01793
.01909
.02027
.02144
.02259
95.37
95.64
95.90
96.17
96.43
96.69
96.95
97.2 1
97.47
97.72
97.98
98.23
98.49
98.74
98.99
99.24
99.48
99.73
99.97
100.22
100.46
100.70
100.94
101.17
101.41
101.64
101.87
102.10
102.33
102.55
102.78
103.00
103.22
103.43
103.65
103.86
104.07
104.28
104.49
104.69
104.89
105.09
105.28
105.47
105.66
0.02376
.02491
.02608
.02722
.02839
0.02954
.03067
.03182
.03296
.03411
0.03526
.03638
.03752
.03865
.03978
0.04092
.04203
.04316
.04430
.04542
0.04654
.04764
.04875
.04986
.05099
0 . 0 5 21 2
.05321
.05431
.05543
.05649
0.05763
.05872
.0597s
.06091
.06203
0 . 0 6 31 1
.06415
.06527
.06637
.06749
0.0685:
.0695$
.07071
.0717S
.07282
IlEMP
(F)
vapor
Ig
0.21421
.21387
.21352
.21318
.21286
0.21252
.21221
.21189
.21160
.21130
0.21101
.21072
.21044
.21017
.20991
0.20965
.20939
.20914
.20890
.20866
0.20843
.20819
.20797
.20775
.20754
0.20732
.20711
.20691
.20672
.2065i
0.2063:
.20614
.2059:
.2057E
.2056(
0.2054;
.2052:
.205OC
.20491
. 2 0 4 7 !i
0 . 2 0 4 5 t1
. 2 0 4 4 :,
. 2 0 4 2 (i
. 2 0 4 1 ()
. 2 0 3 9 !5
0 . 2 0 3 8 (1
. 2 0 3 6 !5
. 2 0 3 5 (1
. 2 0 3 3 :5
. 2 0 3 2 (I
0 . 2 0 3 0 15
.2029I
. 2 0 2 7 ;I
. 2 0 2 6 :2
. 2 0 2 4 13
0 . 2 0 2 3 ~b
. 2 0 2 2 (I
. 2 0 2 0 15
. 2 0 1 9 :2
. 2 0 1 7 13
0 . 2 0 1 6 s4
. 2 0 1 4 19
. 2 0 1 3 15
.20121
. 2 0 1 07:
t
- 4 0
- 3 8
- 3 6
- 3 4
- 3 2
- 3 0
-20
- 2 6
-14
- 2 2
- 2 0
-18
- 1 6
- 1 4
- 1 2
- 1 0
- 8
-6,
4
2
0
2
4
6
II
10
12
14
lb
18
20
22
24
26
28
30
32
34
36
38
40
42
44
46
48
50
52
54
56
58
60
62
64
66
60
70
72
74
76
70
80
82
84
86
88
TABLE 3-PROPERTIES OF REFRIGERANT 500, LIQUID AND SATURATED VAPOR (Contd)
TEMP
(F)
t
90
92
94
96
98
;su
PRES
RE
( l b / ,sq i 4
Abbsolute
Gage
P
P
135.86
121.2
139.83
125.1
143.90
129.2
133.3
148.03
152.27
137.6
iTii
VOI
WE
( C U f t / ll b )
Liquid
Vf
Vapor
r
L
0.0141
.0142
.0142
.0143
.0144
vg
0.3626
.3520
.341a
.3319
.3223
DENSITY
(lb/t ft)
Liquid
l/vt
lI
70.70
70.45
70.19
69.93
69.67
Vapor
11%
2.7581
2.8409
2.9261
3.0128
3.1024
Liquid
hr
36.04
36.66
37.26
37.88
38.47
ENTHALPY
-(Btu/lb)
I 1rrtent
Vapor
69.81
69.37
68.96
68.51
68.10
h,
105.85
106.03
106.22
106.39
106.57
b
l-
ENTROPY
IBtu/lb-R)
Liquid
Vapor
%
51
0.07395
0.20092
.07505
.2007a
.076!1
.20064
.07722
.20049
.20035
.07826
L
I
TEMP
(F)
t
90
92
94
96
98
100
102
104
106
108
156.61
161.02
165.55
170.14
174.84
141.9
146.3
150.9
155.4
160.1
0.0144
.0145
.0145
.0146
.0146
0.3130
.3041
.2953
.2869
.27aa
69.41
69.14
68.87
68.60
68.33
3.1947
3.2889
3.3860
3.4850
3.5871
39.09
39.71
40.33
40.96
41.59
67.65
67.20
66.74
66.27
65.80
106.74
106.91
107.07
107.23
107.39
0.07935
.oao42
.0815-l
.08261
.08367
0.20019
.20004
.I9989
.I9974
.l995a
100
102
104
106
108
110
112
114
116
118
179.62
184.51
189.47
194.55
199.71
164.9
169.8
174.8
179.9
185.0
0.0147
.014a
.oi4a
.0149
.0149
0.2709
.2632
.2558
.2486
.2417
68.05
67.78
67.49
67.21
66.92
3.6914
3.7990
3.9086
4.0218
4.1376
42.22
42.82
43.47
44.11
44.72
65.32
64.87
64.37
63.87
63.40
107.54
107.69
107.84
107.98
108.12
0.08479
.oa482
.08691
.oaaoo
.oa905
0.19942
.I9926
.19910
.I9893
.I9876
110
112
114
116
118
120
122
124
.- .
204.99
210.40
215.88
221.44
222.13
190.3
195.7
201.2
206.7
212.4
0.0150
.0151
.0151
.0152
.0153
0.2349
.22a3
.2219
.2157
.2097
66.63
66.34
66.04
65.74
65.43
4.2566
4.3807
4.5068
4.6357
4.7690
45.35
46.02
46.69
47.33
47.98
62.90
62.35
61.81
61.29
60.75
108.25
108.37
108.50
108.62
108.73
0.09012
.09124
.09236
.09342
.09451
0.19859
.l984i
.I9823
.I9805
.19786
120
122
124
126
128
-i30132
134
136
138
232.89
238.80
244.84
250.96
257.17
218.2
224.1
230.1
236.3
242.5
0.0154
.0154
.0155
.0156
.0157
0.2039
.I982
.I926
.ia72
.1a20
65.13
64.81
64.49
64.17
63.85
4.9050
5.0463
5.1924
5.3420
5.4951
48.65
49.28
49.91
50.59
51.30
60.19
59.66
59.13
58.54
57.91
108.84
108.94
109.04
109.13
109.21
0.09560
.096bb
.09771
.09aa3
.09996
0.19767
.19747
.19726
.19705
.I9684
130
132
134
136
138
140
142
144
146
148
263.52
269.99
276.55
283.24
290.04
248.8
255.3
261.9
268.5
275.3
0.1769
.1719
.1670
.1623
.I577
63.52
63.18
62.84
62.49
62.14
5.6528
5.8170
5.9878
6.1626
6.3431
52.02
52.73
53.45
54.15
54.86
57.27
56.63
55.97
55.33
54.67
109.29
109.36
109.42
109.48
109.53
0.10114
.I0228
.10344
.10456
.I0569
0.19663
.I9639
.19615
.I9591
.I9565
140
142
144
146
148
150
152
154
156
158
296.97
304.01
311.18
318.47
325.87
282.3
289.3
296.5
303.8
311.2
0.0157
.ol58
.0159
.Ol60
.0161
t
0.0162
.0163
.0164
.Ol65
.0166
0.1531
.l4a7
.1444
.1402
.I360
61.78
61.42
61.05
60.66
60.28
6.5304
6.7241
6.9253
7.1342
7.3510
55.62
56.34
57.07
57.82
58.61
53.95
53.25
52.54
51.80
51.01
109.57
109.59
109.61
109.62
109.62
0.10691
.I0806
.10922
.11040
.I1163
0.19539
.19511
.I9483
.19453
.I9421
150
152
154
156
158
160
333.40
318.7
0.0167
0.1320
59.88
7.5772
59.35
50.25
109.60
0.11280
0.19389
160
TABLE 4-PROPERTIES OF REFRIGERANT 22, LIQUID AND SATURATED VAPOR
PRESSURE
(lb/s<
1.)
TEMP
(F)
brolute
P
Gage
VOL
(cu f
l-
\E
b)
Liquid
vt
j.0102
.0103
.OlO3
.OlO3
.0104
vapor
“g
88.1
46.1
14.5
90.61
72.33
DENSITY
(Ib/cu 0)
I
-L
Liquid
Vapor
ILiquid
1 /Vf
hf
l/%3
97.67
1.005316
-29.07
-27.79
97.33
JO6847
96.99
.008733
- 26.52
96.63
.01104
- 25.25
- 23.99
96.27
.Ol383
I
ENTHALPY
( h/lb)
Latent
F
ENTROPY
CBfu/lb-R)
l- EMP
(F)
T
Liquid
Sf
- 0.0808
- .0767
- .0727
- .0687
- .0647
Vapor
I Il.10
110.45
109.80
89.70
90.29
90.88
91.47
92.07
-0.0609
- .0571
- .0534
- .0497
- .0461
0.2803
.2770
.2738
.2708
.2680
-130
-125
-120
-115
-110
92.67
93.27
93.87
94.47
95.08
-0.0425
- .0390
- .0356
- .0322
- .0288
0.2653
.2627
.2602
.2579
.2556
-105
-100
- 95
- 90
- a5
“fg
115.85
115.15
114.46
113.78
113.10
vapor
r
+ ~,
h,
86.78
87.36
87.94
88.53
89.11
sg
0.2996
.2952
.2912
.2874
.2837
I
-155
- 150
-145
-140
-135
0.19901
0.2605
0.3375
0.4332
0.5511
P
29.51*
29.39*
29.23*
29.04*
28.80*
-
0.6949
0.8692
1.079
1.329
1.626
28.51'
28.15*
27.72*
27.21'
26.61'
j.0104
.0105
.OlO5
.0106
.0106
58.21
47.23
38.60
31.77
26.33
95.91
95.53
95.15
94.76
94.37
I.01718
.02118
.02591
.03147
.03798
-
- 105
-100
- 95
- 90
- 85
1.976
2.386
2.845
3.417
A.055
25.90'
25.06*
24.09+
22.96*
21.67*
I.0106
.0107
.0107
.OlO8
.OlOE
21.96
18.43
15.54
13.20
II.26
93.97
93.56
93.14
92.72
92.29
1.04554
.05427
.06433
.07578
.08884
- lb.48
-15.23
- 13.98
- 11.47
109.15
108.50
107.85
107.20
106.55
-
80
78
76
74
72
4.787
5.100
5.430
5.79
6.17
20.18'
19.55*
18.87*
1a.14*
17.37*
).01090
.01091
.01093
.01095
.01097
9.650
9.086
8.561
8.072
7.616
91.85
91.67
91.49
91.31
91.13
L1036
.I101
.I168
.1239
.I313
-10.22
- 9.72
- 9.21
- 8.70
- 8.20
105.90
105.64
105.37
105.10
104.84
95.68
95.92
96.16
96.40
96.64
-0.0255
- .0242
- .0229
- .0216
- .0203
0.2535
.2526
.2518
.2510
.2502
-
-
70
68
66
64
62
6.57
6.99
7.40
7.86
8.35
16.55*
15.70+
14.86*
13.93*
12.93*
3.01100
.01102
.OllO4
.OllOb
.01109
7.192
6.795
6.426
6.079
5.755
90.95
90.77
90.58
90.39
90.21
I.1391
.I472
.I556
.I 6 4 5
.1738
-7.69
- 7.19
- 6.68
- 6.17
- 5.67
104.57
104.31
104.04
103.77
103.51
96.88
97.12
97.36
97.60
97.84
-0.0190
- .0177
- .0164
- .0151
- .Ol38
0.2494
.2487
.2479
.2472
.2465
- 70
- 60
- 66
- 64
- 62
- 60
- 58
- 56
- 54
- 52
8.86
9.39
9.94
10.51
11.11
11.89'
10.81'
9.69'
8.53'
7.31*
0.01111
.01113
.01115
.Oll18
.01120
5.452
5.166
4.900
4.650
4.415
90.03
89.84
89.65
89.46
89.27
I.! 834
.I936
.2041
.2151
.2265
-
5.16
4.65
4.13
3.61
3.09
103.24
102.97
102.69
102.41
102.13
98.08
98.32
98.56
98.80
99.04
-0.0126
- .0113
- .OlOO
- .0087
- .0075
0.2458
.2451
.2444
.2438
.243l
- 60
- 58
- 56
- 54
- 52
- 5c
- 48
- 4c
- 44
- 49
11.74
12.40
13.09
13.80
14.54
6.03*
4.68+
3.28*
1.83'
0.326'
0.01123
.Ol125
.01128
.01130
.01133
4.192
3.986
3.793
3.611
3.440
89.08
88.88
88.68
88.49
88.30
3.2386
.2509
.2636
.2769
.2907
-
2.58
2.06
1.54
1.02
0.51:
101.86
101.58
101.30
101.02
100.74
99.28
99.52
99.76
100.00
100.23
-0.0062
- .0050
- .0037
- .0025
- .OOl2
0.2425
.241&
.2412
.2406
.2400
-
-
4c
3E
36
34
3:
15.31
16.12
16.97
17.85
18.77
0.610
1.42
2.27
3.15
4.07
0.01135
.01138
.Oll40
.01143
.01146
3.279
3.126
2.981
2.844
2.713
88.10
87.90
87.70
87.50
87.29
0.3050
.3199
.3355
.3517
.3686
0.00
0.53
1.05
1.58
2.10
100.46
100.17
99.88
99.59
99.30
100.46
100.70
100.93
101.17
101.40
0.0000
.0013
.0025
.0037
.0050
0.2394
.2389
.2383
.2377
.2372
- 40
- 3c
- 2f
- 2c
- 24
- 2;
19.72
20.71
21.73
22.79
23.88
5.02
6.01
7.03
8.09
9.18
0.01148
.01151
.Oll54
.01156
.01159
2.590
2,474
2.365
2.262
2.165
87.09
86.89
86.69
86.48
86.27
3.3862
.4043
.4229
.4421
.4619
2.62
3.15
3.69
4.22
4.75
99.01
98.71
98.41
98.11
97.81
101.63
101.86
102.10
102.33
102.56
0.0062
.0074
.0086
.0099
.Olll
0.2367
.236l
.2356
.2351
.2346
-
30
28
26
24
22
- 2(
- 18
- 1t
- 1r
- 1:
25.01
26.18
27.39
28.64
29.94
IO.31
11.48
12.69
13.94
15.24
0.01162
.01165
.01168
.01171
.01174
2.074
1.987
1.905
1.827
1.752
0.4822
.5032
.5249
.5474
.5707
5.28
5.82
6.40
6.90
7.43
97.51
97.20
96.89
96.58
96.27
102.79
103.02
103.25
103.48
103.70
0.0123
.0135
.0147
.0159
.0170
0.2341
.2336
.2331
.2326
.2321
-
20
18
16
14
12
-
31.29
32.49
34.14
35.64
37.19
16.59
17.99
19.44
20.94
22.49
0.01 177
.01180
.01183
.01186
.01189
86.06
85.85
85.64
85.43
- 85.21 -
1.681
1.613
1.549
1.488
1.429
84.99
84.78
84.56
84.34
84.12
0.5948
.6198
.6456
.6723
.6997
7.96
8.49
9.02
9.55
10.09
95.96
95.65
95.34
95.03
94.71
103.92
104.14
104.36
104.58
104.80
0.0182
.0194
.0205
.0217
.0228
0.2316
.2312
.2307
.2302
.2298
- 1 0
8
6
4
2
I
38.79
40.43
42.14
43.91
45.74
24.09
25.73
27.44
29.21
31.04
0.01192
.01195
.01198
.01201
.01205
1.373
1.320
1.270
1.221
1.175
83.90
83.68
83.45
83.23
83.01
0.7282
.7574
.7877
.a191
.a514
10.63
11.17
li.70
12.23
12.76
94.39
94.07
93.75
93.43
93.11
105.02
105.24
105.45
105.66
105.87
0.0240
.0251
.0262
.0274
.0285
0.2293
.2289
.2285
.2280
.2276
0
2
4
6
a
l(
1:
11
l(
II
47.63
49.58
51.59
53.66
55.79
32.93
34.88
36.89
38.96
41.09
0.01208
.Ol211
.01215
.01218
.01222
1.130
1.088
1.048
1.009
0.972
82.78
82.55
82.32
82.09
81.86
0.8847
0.9191
0.9545
0.9911
1.029
13.29
13.82
14.36
14.90
15.44
92.79
92.47
92.14
91.81
91.48
106.08
106.29
106.50
106.71
106.92
0.0296
.0307
.0319
.0330
.034l
0.2272
.2268
.2264
.2260
.2257
10
12
14
16
18
21
2:
2'
2(
21
57.98
60.23
62.55
64.94
67.40
43.28
45.53
47.85
50.24
52.70
0.01225
.01229
.01232
.Ol236
.01239
0.936'
0.903'
0.870:
0.839;
0.8101
81.63
81.39
81.16
80.92
80.69
1.067
1.107
1.149
I.191
1.235
15.98
16.52
17.06
17.61
18.17
91.15
90.81
90.47
90.12
89.76
107.13
107.33
107.53
107.73
107.93
0.0352
.0364
.0375
.0379
.0398
0.2253
.2249
.2246
.2242
.2239
20
22
24
26
28
1
1
1
1
1
3
2
2
1
1
0
5
0
5
0
I(
I
t
1
:
(
1
t
*Inches of mercury below one atmosphere.
22.73
2 I .47
20.22
18.98
- 17.73
- 12.73
112.43
i111.76
-155
-150
-145
-140
-135
80
78
76
7 4
72
50
48
46
44
42
- 3 8
- 36
- 34
- 32
.
TABLE 4-PROPERTIES OF REFRIGERANT 22, LIQUID AND SATURATED VAPOR (Contd)
TEMP
(F)
t
30
PRESSURE
(lb/!
in.)
ibrolute
P
Gage
P
55.23
57.83
60.51
63.27
66.11
VOI
(CU
liquid
vt
I.01243
.01247
.01250
.01254
.oi25a
HE
lb)
vapor
"9
0.7816
.7543
.7283
.7032
.6791
Liquid
1 /Vf
80.45
80.21
79.97
79.73
79.49
TY
ft)
vapor
ENTHALPY
(Btu/lb)
Liquid
hf
18.74
19.32
19.90
20.49
21.09
Latent
Vapor
119
I.280
1.326
1.373
1.422
1.473
hf,
89.39
89.01
88.62
88.22
87.81
h,
108.13
108.33
108.52
108.71
108.90
ENTl
PY
(Btu/ '-l b -,R)
Liquid
Vapor
Of
Ig
0.0409
0.2235
.0421
.2232
.2228
.0433
.0445
.2225
.0457
.2222
lEMP
(F)
t
30
32
34
36
38
32
34
36
38
69.93
72.53
75.21
77.97
80.81
40
42
44
46
48
83.72
86.69
89.74
92.88
96.10
69.02
71.99
75.04
78.18
al.40
I.01262
.01266
.01270
.01274
.01278
0.6559
.6339
.6126
.5922
.5726
79.25
79.00
78.76
78.51
78.26
1.525
1.578
1.632
1.689
1.747
21.70
22.29
22.90
23.50
24.1 1
87.39
86.98
86.55
86.13
85.69
109.09
109.27
109.45
109.63
109.80
0.0469
.04ai
.0493
.0505
.0516
0.2218
.2215
.2211
.2208
.2205
40
42
44
46
48
50
52
54
56
58
99.40
102.8
106.2
109.8
113.5
84.70
ea.10
91.5
95.1
98.8
).oi282
.01286
.01290
.01294
.01299
0.5537
.5355
.51a4
,501 A
.A849
78.02
77.77
77.51
77.26
77.01
1.806
1.868
1.929
1.995
2.062
24.73
25.34
25.95
26.58
27.22
85.25
84.80
EA.35
83.89
83.41
109.98
110.14
110.30
110.47
110.63
0.0528
.0540
.0552
.056A
.0576
0.2201
.2198
.2194
.2191
.2iaa
50
52
54
56
58
60
62
4
,6
68
117.2
121.0
124.9
128.9
133;o
102.5
106.3
110.2
114.2
118.3
I.01303
.01307
.01312
.01316
.01320
0.4695
.4546
.4403
.4264
.4129
76.75
76.50
76.24
75.98
75.72
2.130
2.200
2.271
2.346
2.422
27.83
28.46
29.09
29.72
30.35
82.95
82.47
81.99
81.50
81.00
110.78
110.93
1 i 1.08
111.22
111.35
0.0588
.0600
.0612
.0624
.0636
0.2185
.2181
.2178
.2175
.2172
60
62
64
66
68
70
71
74
76
70
137.2
141.5
145.9
150.4
155.0
122.5
126.8
131.2
135.7
140.3
1.01325
.01330
.01334
.01339
.01344
0.4000
.3875
.3754
.3638
.3526
75.46
75.20
74.94
74.68
74.41
2.500
2.581
2.664
2.749
2.836
30.99
31.65
32.29
32.94
33.61
80.50
79.98
79.46
78.94
78.40
111.49
111.63
111.75
1 i 1.88
112.01
0.0648
.066 1
.0673
.0684
.0696
0.2168
.2165
.2162
.2l5a
.2155
70
72
74
76
78
aa
159.7
164.5
169.4
174.5
179.6
3.01349
.01353
.01358
.01363
.oi3ba
0.3417
.3313
.3212
.3113
.3019
74.15
73.89
73.63
73.36
73.09
2.926
3.019
3.113
3.213
3.313
34.27
34.92
35.60
36.28
36.94
77.86
77.32
76.76
76.19
75.63
0.0708
.0720
.0732
.0744
.0756
0.2151
.214a
.2144
.2140
.2137
80
a2
a4
86
aa
9c
94
94
9c
9E
184.8
190.1
195.6
201.2
206.8
145.0
149.8
154.7
159.8
164.9
170.1
175.4
180.9
186.5
192.1
-i12.13
a2
84
86
af
0.2928
.2841
.2755
.2672
.2594
72.81
72.53
72.24
71.95
71.65
3.415
3.520
3.630
3.742
3.855
37.61
38.28
38.97
39.65
40.32
75.06
74.48
73.88
73.28
72.69
-i12.67
12.76
12.85
12.93
I 13.00
0.0768
.0780
.0792
.0803
.0815
0.2133
.2130
.2126
.2122
.2119
90
92
94
96
98
1oc
101
104
lot
1O f
212.6
218.5
224.6
230.7
237.0
197.9
203.8
209.9
216.0
222.3
0.01374
.01379
.oi3a4
.01390
.01396
-.
0.01402
.oi4oa
.01414
.01420
.01426
0.2517
.2443
.2370
.2301
.2233
71.35
71.05
70.74
70.42
70.11
3.973
4.094
4.220
4.347
4.479
40.98
Al.65
42.32
42.98
43.66
72.08
71.47
70.84
70.22
69.58
-i13.06
113.12
113.16
li3.20
113.24
0.0827
.oa39
.0851
.0862
.oa74
0.2115
.2111
.2107
.2lOA
.2100
100
102
104
106
108
11c
11;
114
1lC
‘If
243.4
249.9
256.6
263.4
270.3
228.7
235.2
241.9
248.7
255.6
0.01433
.01440
.01447
.01454
.01461
0.2167
.2104
.2043
.I983
.I926
69.78
69.45
69.12
68.78
68.44
4.614
4.752
4.896
5.043
5.192
44.35
45.04
45.74
46.44
47.14
68.94
68.30
67.64
66.98
66.32
113.29
113.34
I 13.38
113.42
113.46
0.0886
.089a
.0909
.092l
.0933
0.2096
.2093
.2089
.2085
.2081
110
112
114
116
118
. ..LC
12;
124
12c
12f
277.3
284.4
291.7
299.1
306.5
262.6
269.7
277.0
284.4
291.8
0.01469
.01475
.014a3
.01491
.01498
0.1871
.ia25
.I772
.I724
.1675
68.10
67.75
67.40
67.05
66.70
5.345
5.48
5.64
5.80
5.97
47.85
48.60
49.40
50.20
50.80
65.67
64.97
64.21
63.45
62.89
113.52
113.57
113.61
113.65
113.69
0.0945
.0959
.0973
.0986
.0997
0.2078
.2076
.2073
.2070
.2067
120
122
124
126
128
1x
13:
134
13t
13f
314.0
321.8
329.9
338.3
347.0
299.3
307.1
315.2
323.6
332.3
0.01507
.01515
.01523
.01532
.01541
0.1629
.1585
.I538
.1492
.I449
66.35
66.00
65.65
65.25
64.85
6.14
6.31
6.50
6.70
6.90
51.50
52.30
53.10
53.80
54.60
62.21
61.44
60.67
59.99
59.20
113.71
113.74
113.77
113.79
i 13.80
0.1009
.1022
.I035
.1046
.1059
0.2064
.2061
.2057
.2053
.2050
130
132
134
136
138
14c
14:
14r
14f
14f
356.0
365.0
374.1
383.3
392.6
341.3
350.3
359.4
368.6
377.9
0.01551
.01561
.01571
.01581
.01591
0.1408
.I368
.1330
.I292
.1253
64.45
64.05
63.65
63.25
62.85
7.10
7.31
7.52
7.74
7.98
55.30
56.10
56.90
57.70
58.40
58.51
57.70
56.89
56.08
55.36
113.81
113.80
113.79
113.78
113.76
0.1070
.1084
.I096
.I110
.1120
0.2046
.2043
.2039
.2036
.2031
140
142
144
146
148
15(
15:
154
1%
151
401.9
411.3
420.8
430.3
439.8
387.2
396.6
406.1
415.6
425.1
0.01601
.01612
.01625
.01636
.01648
0.1216
.1179
.114l
.I105
.1070
62.45
62.02
61.58
61.13
60.67
8.22
8.48
a.76
9.05
9.35
59.20
60.00
60.80
61.60
62.50
54.54
53.71
52.87
52.02
51.06
113.74
113.71
113.67
113.62
113.56
0.1132
.1145
.I156
.1168
.1181
0.2027
.2023
.2ola
.2013
.2008
150
152
154
156
158
lb(
449.3
434.6
0.01661
0.1035
60.20
9.66
63.50
50.00
113.50
0.1196
0.2003
160
12.24
12.36
12.47
12.57
TABLE 5-PROPERTIES
TEMP
WI
T
T-
PRESSURE
(lb/w 4 i t1.)
,Gage
A bsolufe
P
P
OF REFRIGERANT 1 1, LIQUID AND SATURATED VAPOR
DENSITY
ft)
VOLUME
(cu
b)
(lb/
I .iquid
l/Yf
1 01.25
1 01.10
1 00.96
1 00.8 I
1 00.66
t
ENTHALPY
Btu/ib)
i*
Liquid
hf
0.00
0.40
0.79
1.19
1.59
Latent
Vapor
l/Q
1.02260
.02409
.02566
.02732
.02906
hfe
87.53
87.37
87.22
87.06
86.91
h,
87.53
87.77
88.01
88.25
88.49
liquid
St
0.0000
.0009
.0019
.0028
.0038
SB
0.2086
.2081
.2077
.2073
.2070
00.52
00.37
00.22
00.07
99.92
I.03090
.03282
.03485
.03697
.03920
1.99
2.38
2.78
3.18
3.58
86.75
86.60
86.44
86.29
86.13
Vapor
-40
-38
-36
-34
-32
88.74
88.98
89.22
89.47
89.71
0.0047
.0056
.0065
.0074
.0084
0.2066
.2062
.2058
.2055
.2051
-30
-28
-26
-24
-22
0.7387
0.7911
0.8466
0.9053
0.9674
2!8.42*
2!8.31f
2!8.20*
2!8.08*
i !7.95*
-30
-28
-26
-24
-22
1.0330
I.1023
1.1754
1.253
1.334
2 !7.82*
2 !7.68*
; !7.53*
i !7.37*
i !7.20'
3.00995
.00996
.00998
.00999
.01001
32.36
30.47
28.70
27.05
25.51
-20
-18
- 1 6
- 1 4
- 1 2
1.419
1.509
1.604
1.704
1.809
; !7.031
;!6.85*
:l6.65'
:!6.45*
1 : !6.24*
3.01002
.01004
.01005
.01007
.01008
24.08
22.74
21.48
20.31
19.22
99.78
99.63
99.48
99.33
99.18
I.04154
.04398
.04655
.04923
.05204
3.98
4.38
4.77
5.17
5.57
85.98
85.82
85.67
85.51
85.36
89.95
90.20
90.44
90.69
90.93
0.0093
.0102
.0111
.0120
.0129
0.2048
.2045
.2042
.2038
.2035
-20
-la
-16
-14
-12
- 1 0
-8
-6
- 4
_ 2
1.918
2.034
2.155
2.282
2.415
:l6.01'
:25.78;
>5.53*
:!5.27*
::5.00*
0.01010
.OlOll
.01013
.01014
.01016
18.19
17.23
16.33
15.48
14.69
99.03
98.88
98.72
98.57
98.42
I.05497
.05804
.06123
.06458
.06807
5.98
6.38
6.77
7.19
7.59
85.20
85.05
84.89
84.74
84.58
91.19
91.43
91.67
91.92
92.17
0.0138
.0147
.0155
.0164
.0173
0.2032
.2029
.2027
.2024
.2021
-10
* 8
- 6
- 4
- 2
0
2
4
6
8
2.554
2.699
2.852
3.011
3.178
14.72*
24.42*
24.11 *
23.79*
23.45*
O.OlOd8
.01019
.01021
.01022
.01024
13.95
13.25
12.59
11.97
11.38
98.27
98.11
97.96
97.81
97.65
3.07171
.07550
.07945
.08356
.08785
7.99
8.39
a.79
9.19
9.60
84.43
84.27
84.12
83.96
83.80
92.42
92.67
92.91
93.16
93.40
0.0182
.0191
.019q
.0208
.0217
0.2018
.2016
.2013
.2011
.2009
0
2
4
6
a
10
12
14
16
18
3.352
3.533
3.723
3.921
4.127
23.10*
22.73*
22.34'
21.94*
21.52'
0.01026
.01027
.01029
.01031
.01032
10.83
10.32
9.828
9.367
8.932
97.50
97.35
97.19
97.04
96.88
10.00
10.41
10.81
11.22
11.62
83.65
83.49
83.33
83.18
83.02
93.65
93.90
94.15
94.39
94.64
0.0225
.0234
.0243
.0251
.0260
0.2006
.2004
.2002
.2000
.1998
10
12
14
16
la
20
22
24
26
28
4.342
4.566
4.799
5.041
5.294
21.08'
20.62*
20.15*
19.66*
19.14*
0.01034
.01036
.01037
.01039
.01041
8.521
8.133
7.766
7.418
7.090
96.72
96.57
96.41
96.25
96.10
0.09233
.09694
.1018
.1068
1120
>
0.1174
.1230
.1288
.1348
.1410
12.03
12.43
12.85
13.26
13.67
82.86
82.70
82.55
82.39
82.23
94.89
95.14
95.39
95.65
95.89
0.0268
.0277
.0285
.0294
.0302
0.1996
.1994
.1992
.1990
.1988
20
22
24
26
28
30
32
34
36
38
5.556
5.829
6.112
6.407
6.712
18.61"
18.05'
17.4a*
16.881
16.25*
0.01042
.01044
.01046
.01048
.01049
6.779
6.484
6.205
5.940
5.688
95.94
95.78
95.62
95.46
95.30
0.1475
.1542
.1612
.1684
.1758
14.07
14.48
14.89
15.30
15.71
82.07
81.91
81.75
81.59
81.42
96.14
96.39
96.64
96.89
97.14
0.0310
.0319
.0327
.0335
.0344
0.1986
.1985
.1983
.19Bl
.I980
30
32
34
36
38
40
42
44
46
48
7.030
7.359
7.700
8.054
8.420
15.61'
14.94*
14.24*
13.52'
12.78*
0.01051
.01053
.01055
.01056
.01058
5.450
5.223
5.008
4.804
4.609
95.14
94.98
94.82
94.66
94.50
0.1835
.I914
.1997
.2082
.2169
16.12
16.54
16.95
17.36
17.77
81.26
81.10
80.94
80.77
80.61
97.39
97.63
97.88
98.13
98.38
0.0352
.0360
.0368
.0377
.0385
0.1978
.1977
.1975
.1974
.1973
40
42
44
46
48
50
52
54
56
58
8.800
9.193
9.600
10.02 ,
10.46
12.00*
11.20*
10.37'
9.52'
8.63'
0.01060
.01062
.01064
.01065
.01067
4.425
4.249
4.081
3.922
3.770
94.34
94.18
94.02
93.86
93.69
0.2260
.2354
.2450
.2550
.2653
18.19
18.61
19.03
19.44
19.86
80.44
80.28
80.11
79.94
79.77
98.63
98.89
99.14
99.38
99.63
0.0393
.0401
.0409
.0417
.0425
0.1971
.I970
.1969
.1968
.1966
x52
54
56
58
60
62
64
66
68
10.91
11.37
11.85
12.35
12.86
7.71'
6.76*
5.79*
4.77'
3.73*
0.01069
.01071
.01073
.01075
.01077
3.625
3.487
3.355
3.229
3.109
93.53
93.37
93.20
93.04
92.87
0.2759
.2868
.2981
.3097
.3216
20.27
20.69
21.11
21.53
21.95
79.61
79.44
79.27
79.10
78.92
99.88
100.13
100.38
100.62
100.87
a.0434
.0442
.0450
.0458
.0466
0.1965
.1964
.1963
.1962
.1961
60
62
64
66
68
70
72
74
76
78
13.39
13.94
14.51
15.09
15.69
2.65"
1.54*
0.39'
0.39
0.99
0.01079
.01081
.01083
.01084
:01086
2.995,
2.885
2.781
2.681
2.585
92.71
92.54
92.38
92.21
92.05
0.3339
.3466
.3596
.3730
.3868
22.37
22.79
23.21
23.63
24.06
78.75
78.58
78.41
78.23
78.05
101.12
101.37
101.61
101.86
102.12
0.0474
.0481
.0489
.0497
'.0505
0.1960
.1959
.1959
.I958
.1957
70
72
74
76
78
a0
a2
a4
86
aa
16.31
16.94
17.60
18.28
18.97
1.61
2.25
2.91
3.58
4.28
0.01088
.01090
.01092
.01094
.01096
2.494
2.406
2.322
2.242
2.165
91.88
91.71
91.54
91.38
91.21
0.4010
.4156
.4306
.4461
.4619
24.48
24.91
25.33
25.76
26.18
77.88
77.70
77.52
77.34
77.1 6
102.36
102.61
102.85
103.10
103.34
0.0513
.0521
.0529
.0537
.0544
0.1956
.1955
.1955
.1954
.1953
a0
a2
a4
86
aa
90
19.69
20.43
21.19
21.97
22.77
4.99
5.73
6.49
7.27
8.08
0.01098
.01100
.01103
.01105
.01107
2.091
2.021
1.953
1.888
1.826
91.04
90.87
90.70
90.53
90.36
0.4781
.4949
.5121
.5297
.5477
26.61
27.03
27.46
27.89
28.32
76.98
76.80
76.62
76.43
76.25
103.59
103.83
104.08
104.32
104.56
0.0552
.0560
.056B
.0575
.0583
0.1953
.1952
.1951
.1951
.I950
90
92
94
96
98
-40
-30
-36
-34
-32
;;
96
98
*Inches
Yg
44.25
41.51
38.97
36.60
34.41
1
1
1
1
of mercury below one atmosphere.
,
vapor
T EMP
r (F)
ENTROPY
(Bfu/lb-R)
liquid
vi
3.00988
.00989
.00991
.00992
.00993
t
vapor
r
t
.
~ ; 1 1 . \ 1 ” 1 ’ 1 ~ 1 <I.
1~1:l~I~I~~I~:I~.\N’I’S
TABLE ii--PROPERTIES
lEMs I
(F)
t
-ii%
PRESSURE
W/r ;-q7i n.1
Absolute
P
Gage
r-
417
OF REFRIGERANT 1 1, LIQUID AND SATURATED VAPOR (Contd)
V O Liii A E
( c u f ‘l/l l b )
liquid
Vf
0.01109
102
104
106
108
23.60
24.45
25.32
26.21
27.13
P
8.90
9.75
10.62
11.52
12.44
110
112
114
116
118
28.08
29.05
30.04
31.06
32.11
13.38
14.35
15.35
16.37
17.41
0.01119
.01122
.01124
.01126
.Oll28
120
122
124
126
128
33.18
34.29
35.42
36.57
37.76
18.49
19.59
20.72
21.88
23.06
0.01130
.01133
130
132
134
-16
8
38.97
40.22
41.49
42.80
44.13
24.28
25.52
26.80
28.10
29.44
140
142
144
146
148
45.50
46.90
48.33
49.80
51.29
30.81
32.20
33.64
35.10
36.60
0.01142
.Olldd
.01146
.01149
.01151
0.01154
.01156
.0115B
150
152
154
156
158
52.83
54.39
55.99
57.63
59.30
38.13
39.71
41.31
42.94
44.61
0.01166
.01168
.01171
.01173
.Ol.l76
160
61.01
46.31
0.01178
.Ollll
.01113
.01115
.01117
.01135
.01137
.01139
.01161
.01163
r
DE?
(lb1
1.364
1.322
1.282
1.243
1.206
1.170
1.135
1.101
1.069
1.038
1.008
0.9788
liquid
l/W
90.19
90.02
89.85
89.68
89.51
89.34
89.16
88.99
88.82
88.65
88.47
88.30
88.12
87.95
87.77
87.60
87.42
87.25
87.07
86.89
0.9508
.923B
‘8977
.8725
.84B2
86.68
86.50
86.32
86.14
85.96
0.8247
JO20
.7800
.7588
.7382
0.7183
85.78
85.60
85.41
85.23
85.04
84.86
vapor
VB
1.766
1.708
1.653
1.600
1.549
1.500
1.453
1.408
TY
ft)
Vapor
-r
ENTHALPY
1 /%a
liquid
hf
0.5663
.5854
.6049
.6250
.6455
0.6667
.6882
.7104
.7330
.7563
0.7801
.8045
.8295
.8551
.8812
0.9081
0.9355
0.9636
0.9923
1.022
1.052
1.082
1.114
1.146
1.179
1.213
1.247
1.282
1.318
1.355
1.392
28.75
29.17
29.62
30.05
30.48
30.91
31.34
31.78
32.21
32.65
33.08
33.52
33.95
34.40
34.84
35.28
35.72
36.16
36.60
37.04
37.48
37.92
38.37
38.81
39.25
39.70
40.15
40.60
41.05
41.50
41.95
ENTROPY
(Em/lb-R)
TEMP
(F)
100
102
104
106
108
110
112
114
116
118
73.34
108.18
73.14
72.94
72.73
72.52
72.31
108.42
108.65
71.47
71.25
108.89
109.12
109.35
110.28
110.51
110.74
1 10.96
111.20
111.43
.I945
.1945
0.0704
.0712
.0719
.0726
.0733
0.1945
.1944
0.0741
0.1944
.1943
.1943
.1943
140
142
144
146
148
0.0778
.0785
.0792
0.1943
.OBOO
.1943
150
152
154
156
158
.1945
.1945
.1944
.1944
.1944
--T.0749
.0756
.0763
.0770
.1943
-T- --I71.04
70.82
70.60
70.38
120
122
124
126
128
.0674
.0682
.0689
.0697
.1943
.1943
130
132
134
136
138
160
TABLE 6-PROPERTIES
TEMP
(F)
t
T-
l-
PRESSURE
(lb/w q it1.)
,Gage
A bsolute
P
P
I
OF REFRIGERANT 113, LIQUID AND SATURATED VAPOR
VOLUME
(cu f’
liquid
4Clp.W
vg
32.26
'6.8 1
'1.71
56.99
52.63
1
1
1
1
1
1
l-
DENSITY
( l b / <:” ft)
L
iquid
Vapor
I.iquid
1 /Vf
ht
11%
05.64
0.01216
1.97
05.50
.01302
2.36
05.37
.01395
2.76
05.23
.01493
3.16
3.56
05.09
.01597
0.1732
.I731
.1730
.1729
.1729
-20
-1.3
-16
-14
-12
0.0137
.0146
.0155
.0164
.0173
0.1728
.1727
.1726
.I726
.I725
-10
- a
-6
- 4
- 2
0.0182
.0190
.0199
.0208
.0216
0.1725
.I724
.I724
.1723
.1723
0
2
4
6
a
10.00
10.41
10.81
11.22
11.62
0.0225
.0234
.0242
.0251
.0259
0.1723
.1722
.1722
.I722
.1722
10
12
14
16
18
12.03
12.44
12.85
13.26
0.0268
.0276
.06929
13.67
.0285
.0293
.0302
0.1722
.1721
.1721
.1722
.1722
20
22
24
26
28
0.07294
.07675
.08071
.08483
.08913
14.08
14.49
14.91
15.32
15.74
0.0310
.0318
.0327
.0335
.0343
0.1722,
.1722
.1722
.1722
.1722
30
32
34
36
38
.09361
.09826
.I031
.1081
.1133
16.16
lb.57
0.0352
0.1723
.I723
.1723
.1724
.1724
40
42
44
46
48
0.1187
.1243
.1302
.1362
.1425
18.24
18.66
19.08
19.50
19.93
0.0393
.0401
.0410
.0418
.0426
0.1725
.1726
.1726
.1727
.I727
50
52
54
56
58
5.889
5.640
99.05
98.89
98.73
98.58
98.42
0.1490
.1557
.I626
.1698
.I773
20.35
20.77
21.19
21.62
22.05
0.0434
.0442
.0450
.0459
.0467
0.1728
.I729
.I729
.1730
.1731
60
62
64
66
68
0.01018
.01019
.01021
.01023
.01025
5.404
5.180
4.971
4.769
4.574
98.26
98.10
97.93
97.77
97.6 1
0.1851
.I931
.2012
.2097
22.48
22.90
23.33
0.0475
.0483
.0491
.0499
.0507
0.1731
.1732
.1733
.1734
.1735
70
72
74
76
78
15.87*
15.25*
14.60*
13.93:
13.24*
0.01026
0.1028
.01030
.01031
.01033
4.392
4.218
4.05 1
3.893
3.742
97.45
97.28
97.12
0.2277
.2371
0.0515
.0523
.0531
.oi39
.0547
0.1736
.1737
.1738
.1739
.I740
a0
a2
a4
86
aa
0.01035
.01037
.01039
.01040
.01042
3.600
3.463
3.333
3.208
3.089
96.63
96.46
96.30
96.13
95.96
0.2778
.2a88
, 3 0 01
.3117
.3237
26.80
27.24
10.07
12.53*
11.79'
11.03*
10.24'
9.42*
0.0555
.0563
.0571
.0578
.0586
0.1741
.1742
.I743
.1744
.1745
90
92
94
96
98
10.48
10.91
11.35
11.81
12.28
a.591
7.71'
6.82'
5.88*
4.93:
0.01044
.01046
.01048
.01050
.01051
2.976
2.867
2.762
2.662
2.567
95.79
0.3360
.3488
.3620
28.99
29.44
29.89
30.33
30.78
0.0594
.0602
.0610
.0618
0.1746
.1747
.1748
.1750
.17.51
100
102
104
106
108
0.4288
0.4600
0.4931
0.5280
0.5652
;! 9 . 0 5 *
;! 8 . 9 8 *
;! 8 . 9 2 "
;! 8 . 8 5 *
;! 8 . 7 7 "
1.00953
.00954
.00955
.00957
.00958
58.61
54.88
51.42
18.23
15.25
1 04.96
I 04.82
I 04.68
1 04.54
1 04.40
0.01706
.01822
.01945
.02074
.02210
3.96
4.36
0.6046
3.00959
.00960
.00962
.00963
.00964
42.48
39.92
37.54
35.31
33.24
I 04.26
1, 0 4 . 1 2
11 0 3 . 9 8
11 0 3 . 8 4
11 0 3 . 7 0
0.02354
.02505
0.7369
0.7860
;! 8 . 6 9 *
!8.60*
18.51'
18.42;
!a.32*
.02a32
.03009
5.96
6.36
6.76
7.17
7.57
0.8377
0.8924
0.9503
1.011
1.075
1a.21*
28.10*
27.99'
?7.86*
27.73'
0.00966
.00967
.00968
.00970
.00971
31.31
29.52
27.84
26.27
24.81
11 0 3 . 5 6
103.41
103.27
11 0 3 . 1 3
11 0 2 . 9 8
0.03 194
.03388
.03592
.03806
.04031
7.98
8.38
8.78
9.19
9.59
10
12
14
16
18
1.142
1.213
1.280
1.366
1.448
27.60'
27.45'
27.30'
27.14*
26.97+
0.00972
.00974
.00975
.00977
.00978
23.45
22.17
20.97
19.84
18.79
102.84
102.69
102.55
102.40
102.25
0.04265
.04511
.04769
.05040
.05322
20
22
24
26
28
1.534
1.624
1.719
1.818
1.922
26.80'
26.61;
26.42'
26.22*
26.01*
0.00979
.00981
.00982
.00984
.00985
17.81
16.89
0.05616
lb.02
15.20
14.43
102.10
101.96
101.81
101.66
101.51
30
32
34
36
38
2.031
2.145
2.264
2.388
2.519
25.79'
25.55'
25.31'
25.06'
24.79+
0.00987
.00988
.00990
.00991
.00993
13.71
13.03
12.39
11.79
11.22
101.36
101.21
101.06
100.91
100.76
40
42
44
46
48
2.655
24.52"
24.23*
23.93'
0.00994
100.60
50
52
54
56
58
3.427
3.602
3.784
3.973
4.170
2-3 . 2 9-*
22.94,
22.59'
22.22+
21.83'
21.43*
.00997
.00999
.01000
10.68
10.18
9.703
9.253
8.830
0.0 1002
.01003
.01005
.01006
.01008
8.426
8.044
7.682
7.342
7.018
99.83
99.68
99.52
99.37
99.21
60
62
64
66
68
4.374
4.586
4.807
5.036
5.275
21.02*
20.59*
20.14'
19.67"
19.18'
0.01010
.OlOll
.01013
.01015
.01016
6.713
70
72
74
76
78
5.523
5.780
6.042
18.68*
a0
a2
a4
86
aa
6.902
90
92
94
96
98
a.545
8.908
9.281
100
102
104
106
ioa
0
2
4
6
a
vapor
0.0092
.OlOl
.OllO
.0119
.0128
1 !9.31'
1! 9 . 2 7 *
2! 9 . 2 2 *
2! 9 . 1 6 *
:! 9 . 1 1 *
- 1 0
- a
- 6
- 4
- 2
liquid
Sf
0.0047
.EMP
W
t
-30
.0065
.0074
.0083
0.2987
0.3214
0.3458
0.3718
0.3995
-20
--la
-16
-14
-12
ENTROPY
mu/ lb -R)
Ig
0.1738
.1737
.I736
.1735
.1733
v(
1.00947
.00948
.00949
.00950
.00952
-30
-28
-26
-24
-22
ENTHALPY
(I h/lb)
0.6462
0.6902
2.797
2.944
3.098
3.258
6.320
6.607
7.208
7.527
7.856
8.194
9.660
23.61*
18.16f
17.62*
17.06'
16.47'
.00996
6.424
6.149
100.45
100.30
100.14
99.99
96.96
96.79
95.63
95.46
95.29
95.12
.02664
.05922
.06243
.06579
.2186
.2468
.2569
2672
.3756
.3896
.0056
72.33
1 75.49
4.76
5.16
5.56
71.39
71.27
70.80
70.68
70.56
77.75
78.03
79.18
79.46
79.75
.0360
16.99
17.41
17.82
.0368
.0377
.0385
23.76
24.19
24.63
25.06
25.49
25.93
26.36
27.67
28.11
28.55
65.32
91.68
65.18
65.04
64.90
64.75
91.98
92.28
92.57
92.86
.0626
Cmrrtrsy
*Inches of mercury below one atmosphere.
.
-28
-26
-24
-22
'
TABLE 6--PROPERTIES
TEMP
(F)
t
-Xii
112
114
116
118
120
122
124
126
128
130
132
134
136
138
140
142
144
7°C
-12”
152
154
1Sb
158
160
PRESSURE
(Ib/sq in.)
Absolute 1 Gage
-iTG-k
18.45
19.11
19.79
20.48
(
3.74
4.41
5.09
5.78
l-
OF REFRIGERANT 113, LIQUID
VOLUME
(CU
Liquid
vt
0.01053
.01055
.01057
.01059
.OlObl
0.01063
.01065
.01067
.01069
.01071
0.01073
.01075
.01077
.01079
.01081
0.01083
.01085
.01087
.01089
.01092
0.0109A
.01096
.01098
.01100
.01102
0.01105
b)
Vapor
vg
2.477
2.391
2.308
2.228
2.151
2.078
2.008
1.941
1.876
1.814
1.754
1.697
1.642
1.590
1.540
1.491
1.444
1.399
1.355
1.313
1.273
1.234
1.197
1.162
1.128
1.094
T-
DENSITY
(lb. /cu ft)
Liquid
Vapor
l/V?
l/%
94.95
0.4038
94.78
.4182
94.61
.4333
94.43
.4489
94.26
.4649
94.09
0.4813
93.92
.4981
93.74
.5153
93.57
.5330
93.39
.5514
93.22
0.5702
93.04
.5894
92.86
.6091
92.69
.6290
92.51
.6494
92.33
0.6707
92.15
.6926
91.98
.7150
91.80
.7379
91.62
.7615
91.44
0.7856
91.25
.8102
91.07
.8353
90.89
.a608
90.71
.a609
90.53
0.9141
r
AND SATURATED VAPOR (Contd)
ENTHALPY
Liquid
hf
31.22
31.67
32.12
32.57
33.03
33.48
33.93
34.38
34.83
35.29
35.75
36.21
36.67
37.13
37.59
38.05
38.52
38.98
39.45
39.92
40.38
40.85
41.32
41.79
42.26
42.74
-(Btu/lb)
Latent
hfe
63.71
63.56
63.40
63.25
63.09
62.93
62.78
62.62
62.46
62.30
62.14
61.97
61.80
61.64
61.48
61.31
61.13
60.96
60.79
60.6 1
60.44
60.27
60.09
59.91
59.73
59.55
ENTROPY
Vapor
hg
94.93
95.23
95.52
95.82
96.12
96.4 1
96.71
97.00
97.29
97.59
97.89
98.18
98.47
98.77
99.06
99.36
99.65
99.94
100.24
100.53
100.82
101.11
101.41
101.70
101.99
102.29
Sf
0.0634
.0641
.0649
.0657
.0665
0.0673
.0680
.0688
.0696
.0704
0.0712
.0719
.0727
.0735
.0742
0.0750
.0758
.0765
.0773
.0781
0.0789
.0796
.0804
.0812
.0819
0.0827
vapor
Ig
0.1752
.1753
.I755
.1756
.I757
0.1758
.1760
.I761
.1763
.I764
0.1765
.I767
.1768
.1770
.I771
0.1773
.I774
.1775
.I777
.I778
0.1780
.1782
.1783
.1785
.I786
0.1788
l- TEMP
(F)
t
110
112
114
116
118
120
122
124
126
128
130
132
134
136
138
140
142
144
146
148
150
152
154
156
158
160
TABLE 7-PROPERTIES
TEMP
(F)
t
PRES
(lb/s
,brolute
P
RE
n.)
GLlge
l-
OF REFRIGERANT 114, LIQUID
VOLUME
(cu
lb)
liquid
vt
3.009469
.009484
.009506
.009513
.009528
“cl
j1.26
19.83
14.87
11.69
18.92
Vapor
r
DENSITY
(lb/
ff)
L
liquid
Vapor
1 lVf
l/v,
105.603
0.01951
105.398
.02007
.02229
105.193
105.058
.0240
104.924
.0257
l-
AND SATURATED VAPOR
ENTHALPY
liquid
ht
-8.73
-8.30
-7.87
-7.44
-7 . 0 1
Latent
hfa
ENTROPY
vapor
f
69.17
69.01
68.85
68.69
68.53
h,
60.44
60.71
60.98
61.25
61.52
mu/
liquid
If
-R)
l-
Vapor
TEMP
(F)
t
- 80
-70
-76
-74
-72
0.464
0.50
0.535
0.58
0.62
P.
!8.97"
!8.90*
!8.85*
!8.73*
18.66'
-70
-68
-66
-64
-62
0.670
0.72
0.775
0.833
0.895
!8.59*
!8.46*
!8.33*
!8.23*
!8.10*
3.009543
.009558
.009573
.009589
.009604
$6.40
34.20
$1.74
19.77
27.79
104.790
104.624
104.458
104.292
104.126
0.02747
.02925
.0315
.0336
.0360
-6.57
-6.14
-5 . 7 1
-5.28
-4.84
68.36
68.20
68.04
67.88
67.72
61.79
62.06
62.33
62.60
62.88
-60
-58
-56
-54
-52
0.959
1.028
1.10
1.175
1.26
27.99;
!7.83*
t7.68'
!7.52*
t7.35*
0.009619
.009635
.009651
.009666
.009682
26.06
24.55
22.94
21.50
20.16
103.960
103.792
103.622
103.452
103.282
0.03838
.04075
.0436
.0465
.0496
-4.40
- 3.96
-3.52
-3.08
-2.64
67.56
67.39
67.22
67.05
66.89
63.16
63.43
63.70
63.97
64.25
-50
-48
-46
-44
-42
1.349
1.438
1.535
1.635
1.745
17.20'
17.0 *
16.8 *
26.6 *
26.4 *
0.009698
.009714
.009731
.009747
.009764
18.96
17.85
16.80
15.75
14.87
103.113
102.938
102.766
102.594
102.422
0.05274
.0560
.0595
.0635
.06725
-2.20
-1.76
-1.32
-0.88
-0.44
66.73
66.56
66.39
66.23
66.07
64.53
64.80
65.07
65.35
65.63
-0.0054
- .0043
- .0032
- .0021
- .OOlO
0.1575
.I574
.1573
.1572
.I571
-50
-48
-46
-44
-42
-40
-38
-36
-34
-32
1.866
1.990
2.121
2.259
2.404
26.12*
25.87*
25.60*
25.32"
25.03'
0.00978
.00980
.00981
.00983
.00985
14.02
13.20
12.44
11.73
11.07
102.25
102.08
101.90
101.72
101.55
0.07132
.07574
.08038
.08524
.09034
0.00
0.45
0.91
1.36
1.81
65.91
65.74
65.56
65.38
65.21
65.91
66.19
66.47
66.74
67.02
0.0000
.OOll
.0021
.0032
.0042
0.1571
.1570
.1569
.1568
.I567
-40
-38
-36
-34
-32
-30
-28
-26
-24
-22
2,557
2.718
2.887
3.064
3.249
24.72*
24.39"
24.04*
23.68'
23.31'
0.00987
.00988
.00990
.00992
.00994
10.45
9.877
9.338
8,.833
8.362
101.37
101.19
101.01
100.83
100.65
0.09568
.1013
.1071
.1132
.1196
2.27
2.72
3.17
3.63
4.08
65.03
64.86
64.68
64.51
64.34
67.30
67.58
67.86
68.14
68.42
0.0053
.0063
.0074
.0084
.0095
0.1567
.I567
.1566
.1565
.1565
-30
-20
-26
-24
-22
-20
-18
-16
-14
-12
3.444
3.648
3.862
4.085
4.319
22.91'
22.49*
22.06*
21.61*
21.13*
0.00995
.00997
.00999
.OlOOl
.01003
7.921
7.508
7.121
6.757
6.416
100:47
100.29
100.11
99.92
99.74
0.1263
.I332
.I404
.1480
.1559
4.54
4.99
5.44
5.90.
6.35
64.16
63.99
63.81
63.64
63.46
68.70
68.98
69.26
69.54
69.82
0.0105
.0116
.0126
.0136
.0146
0.156;
.1565
.1564
.I564
.I564
-20
-18
- 1 6
- 1 4
- 1 2
-10
- 8
- 6
- 4
- 2
4.564
4.819
5.086
5.365
5.655
20.63'
20.11*
19.57'
19.00'
18.41*
0.01005
.01006
.01008
.01010
.01012
6.095
5.794
5.510
5.244
4.992
99.56
99.37
99.19
99.00
98.81
0.1641
.1726
.I815
.1907
.2003
6.81
7.26
7.72
8.18
8.63
63.29
63.11
62.94
62.77
62.59
70.10
70.38
70.66
70.94
71.22
0.0157
.0167
.0177
.0187
.0197
0.1564
.1564
.1564
.1564
.I565
- 1 0
-8
- 6
- 4
- 2
0
5.958
6.274
6.603
6.945
7.301
17.79*
17.15"
16.48'
15.78*
15.06*
0.01014
.01016
.01018
.01020
.01022
4.756
4.533
4.322
4.123
3.935
98.62'
98.44
98.25
98.06
97.87
0.2103
.2206
.2314
.2425
.2541
9.09
9.54
10.00
10.46
IO.91
62.42
62.24
62.07
61.89
61.71
71.50
71.78
72.07
72.35
72.63
0.0207
.0217
.0227
.0236
.0246
0.1565
.1565
.1565
.1566
.1566
0
2
4
6
0
15
16
18
7.671
8.057
8.457
8.873
9.305
14.31*
13.52'
12.7! *
11.86*
10.98s
0.01024
.01026
.01028
.01030
.01032
3.758
3.59J
3.432
3.282
3.140
97.68
97.48
97.29
97.10
96.90
0.2661
.2785
.2914
.3047
.3185
11.37
11.83
12.29
12.75
13.20
61.54
61.36
61.19
61.01
60.83
72.91
73.19
73.47
73.75
74.04
0.0256
.0266
.0275
.0285
.0295
0.1566
.1567
.1567
.I568
.1568
10
12
14
16
18
20
22
24
26
28
9.753
10.22
10.70
II.20
11.72
10.07*
9.12'
a.14*
7.12'
6.07*
0.01034
.01036
.01038
.01040
.01043
3.005
2.877
2.756
2.641
2.532
96.71
96.51
96.32
96.12
95.92
0.3328
.3476
.3629
.3786
.3949
13.66
14.12
14.58
15.05
15.51
60.65
60.48
60.30
60.12
59.94
74.32
74.60
74.88
75.17
75.45
0.0304
.0314
.0323
.0333
.0342
0.1569
.1569
.1570
.I571
.1571
20
22
24
26
28
30
32
34
36
38
12.25
12.81
13.38
13.98
14.59
4.99'
3.05*
2.69'
1.47*
0!221
0.01045
.01047
.01049
.01051
.01053
2.429
2.330
2.236
2.147
2.062
95.73
95.53
95.33
95.13
94.93
0.4118
.4292
.4472
.4658
.4849
15.97
16.43’
16.89
17.36
17.82
59.76
59.58
59.40
59.22
59.04
75.73
76.01
76.29
76.58
76.86
0.0352
.036l
.0370
.0380
.0389
0.1572
.1573
.I574
.1575
.1575
30
32
32
36
38
40
42
44
46
48
15.22
15.88
16.56
17.26
17.98
0.52
1.18
1.86
2.56
3.28
0.01056
.01058
.01060
.01063
.01065
1.982
1.905
1.832
1.762
1.695
94.73
94.52
94.32
94.12
93.91
0.5047
.5250
.5460
.5676
.5899
18.28
18.75
19.21
19.68
20.14
58.86
58.67
58.49
58.31
58.12
77.14
77.42
77.70
77.99
78.27
0.0398
.0408
.0417
.0426
.0435
0.1576
.I577
.1578
.I579
.1580
40
42
44
46
48
50
52
54
56
58
18.73
19.50
20.29
21.11
21.96
4.03
4.80
5.59
6.41
7.26
0.01067
.01070
.01072
.01074
.01077
1.632
1.571
1.513
1.458
1.405
93.71
93.50
93.30
93.09
92.88
0.6129
.6365
.6609
.6859
.7117
20.61
21.08
21.54
22.01
22.48
57.94
57.75
57.56
57.38
57.19
78.55
78.83
79.11
79.39
79.67
0.0444
.0453
.0463
90472
.0481
0.1581
.1582
.1583
.1584
.1585
50
52
54
56
58
2
4
6
a
10
12
*Inches of mercury below one atmosphere.
-0.0227
- .0215
- .0203
- .Ol91
- .0179
Ig
0.1595
.I593
.1592
.1590
.1589
- 80
-78
-76
-74
-72
-0.0167
- .0155
- .Ol43
- .0132
- . 0121
~
0.1587
.1586
.1585
.I583
.1582
-70
-60
-66
-64
-62
-0.0110
.-. .0098
- .0087
- .0075
- .0064
0.1580
.1579
.1578
.1577
.l576
-60
-58
-56
-54
-52
l
(:I
I \I’1
I~.I<
I,
I~I~:l~‘I~I(;~:l~.\s’I’s
4-2 1
TABLE ‘/-PROPERTIES OF REFRIGERANT 114, LIQUID AND SATURATED VAPOR (Conic!)
TEMl
W
I
60
62
64
66
68
70
72
74
76
78
PRESSURE
(lb/w
.brolute
P
22.83
23.72
24.64
25.59
26.57
P
a.13
9.02
9.94
lo.89
II.87
27.57
28.61
29.67
12.87
13.91
14.97
1.354
1.306
1.260
1.216
1.174
0.9869
0.01104
.01107
.OlllO
.01112
.01115
0.9541
90.56
.a353
90.34
90.13
89.91
89.69
42.02
43.44
44:a9
24.59
25.94
27.32
28.74
30.19
0.01 i ia
.01120
.01123
.01126
.01129
0.8084
.7a27
.7579
.7340
.7111
46.39
31.69
47.92
49.48
51.09
52.73
33.22
34.78
0.01132
.01135
.01137
.01140
.01143
0.6890
A677
54.41
39.71
41.44
43.20
45.00
46.85
0.01146
48.74
0.01162
52.65
.01165
.oliba
30.76
31.88
90
92
94
-_
39.29
120
122
124
126
128
Vapor
vg
36.69
37.97
40.64
56.14
57.90
59.70
61.55
63.44
65.37
67.35
1.133
1.094
1.057
1.021
17.18
0.01091
.01094
.01097
.01099
.01102
icr
19.52
20.74
21.99
23.27
DENSITY
(lb/
ff)
Liquid
Vapor
1 IVf
l/h
92.68
0.7383
92.47
.7655
92.26
.7936
92.05
.a225
91.84
.a521
0.8826
0.9140
33.04
34.22
35.44
110
112
114
116
ila
Liquid
Vf
0.01079
.oioa2
.oioa4
.01086
.oioa9
HE
lb)
91.63
91.41
91.20
90.99
90.77
a0
02
a4
06
aa
-100
102
104
106
108
Gage
vo
(CU
16.06
36.39
38.03
.01149
.01152
.OllSb
.01159
0.9462
0.9793
1.013
Liquid
hf
22.95
23.42
23.89
24.36
24.83
1atem
56.43
56.24
h,
79.95
80.23
80.51
80.79
al.07
25.30
25.78
56.04
al.35
55.65
al.90
82.18
82.46
26.25
26.73
hfg
57.00
56.81
56.62
55.85
55.45
27.20
55.26
1.048
I.084
1.121
1.159
1.197
27.68
28.15
28.63
29.11
29.58
55.06
89.47
89.25
89.03
88.81
88.59
1.237
1.278
1.320
1.406
30.06
30.54
31.02
31.50
31.99
.boa4
88.37
88.15
87.93
87.70
87.48
1.452
1.498
1.545
1.594
1.644
0.5901
.5724
.5554
.53a9
.5230
87.25
87.01
86.77
86.54
86.31
0.5077
.4929
.47a7
.9226
.a923
.a632
.6472
.6274
ENTROPY
( B t u / lbl R)
ENTHALPY
h/lb)
V C!por
81.62
liquid
Sf
0.0490
.0499
.05oa
.0517
.0526
0.0534
.0543
.0552
.0561
.0570
54.25
82.73
83.01
83.29
83.56
83.84
0.0579
.05a7
.0596
.0605
.Obl3
54.05
53.84
84.1 i
84.39
0.0622
53.64
84.66
32.47
32.95
33.43
33.92
34.40
53.01
52.80
52.59
52.37
85.48
85.75
86.02
1.695
1.747
1.801
i.a56
1.912
34.89
35.38
35.87
1.362
54.86
54.66
54.46
53.43
53.22
84.93
85.21
.063l
.0639
.0648
.0656
0.0665
.0674
.0682
‘EMP
(F)
vapor
60
62
64
66
68
0.1593
.I594
.1595
.1596
.1597
70
72
74
76
78
0.1599
i .I600
a0
a2
a4
86
aa
.I601
.1603
*lb04
0.1605
.1607
.I608
.1609
.lbll
90
92
94
96
98
0.1612
.I614
.1615
.1617
.1618
100
102
104
106
108
110
112
114
116
11.9
86.29
86.56
.0691
86.83
87.09
87.36
87.89
0.0708
.0716
.0725
.0733
.0741
0.1619
.1621
.1622
36.35
36.84
51.94
51.72
51.50
51.28
51.05
1.970
2.029
2.089
2.151
2.215
37.33
37.83
38.32
38.81
39.31
50.83
50.60
50.37
50.14
49.91
88.16
88.42
88.69
88.95
89.21
0.0750
.075a
.0767
.0775
.0783
0.1627
0.1634
.1635
52.16
87.63
.0699
t
)B
0.1587
.i5aa
.15a9
.1590
.1591
.1624
.1625
.01171
.01175
.4649
.4515
86.08
as.85
85.61
85.37
85.13
0.01178
.oiiai
.oi185
.oiiaa
.01192
0.4387
.4262
.4142
.4025
.3912
84.89
84.65
84.41
84.16
83.91
2.280
2.346
.2.414
2.484
2,556
39.80
40.30
40.80
41.29
41.79
49.67
77.90
80.15
82.44
60.99
63.20
65.45
67.74
49.44
49.20
48.96
48.72
89.47
89.73
89.99
90.25
90.51
0.0792
.oaoo
.oaoa
.0816
.oa25
.1637
.163a
.1640
130
132
134
136
138
84.79
70.09
0.01195
0.3803
83.66
2.629
42.29
48.47
90.76
0.0833
0.1641
140
50.67
69.37
71.43
54.67
130
132
134
136
138
73.54
58.84
75.69
140
56.73
.1628
.1630
.1631
.I633
120
122
124
126
128
4-B
CHAPTER 2. BRINES
This chapter provides information to guide the
engineer in the selection of brines, and includes
the properties of the commonly used brines.
At temperatures above 32 F, water is the most
commonly used heat transfer medium for conveying
a refrigeration load to an evaporator. At temperatures below 32 F, brines are used. They may be:
1. An aqueous solution of inorganic salts, i.e.
sodium chloride or calcium chloride. For low
temperatures, a eutectic mixture may be used.
An aqueous solution of organic compounds,
i.e. alcohols or glycols. Ethanol water, methanol water, ethylene glycol and propylene
glycol are examples.
3. Chlorinated or fluorinated hydrocarbons and
halocarbons.
A solution of any salt in water, or in general any
solution, has a certain concentration at which the
freezing point is at a minimum. A solution of such
a concentration is called a eutectic mixture. The
temperature at which it freezes is the eutectic temperature. A solution at any other concentration
starts to freeze at a higher temperature. Figure 11
illustrates the relationship between the freezing
point (temperature) of a brine mixture and the
percent of solute in the mixture (concentration).
Chart 18 covers a range of temperatures wide
enough to reveal the two freezing point curves.
-“hen the temperature of a brine with a concentrr_-,on below the eutectic falls below the freezing
point, ice crystals form and the concentration of
the residual solution increases until at the eutectic
temperature the remaining solution reaches a eutectic concentration. Below this temperature the mix-’
ture solidifies to form a mechanical mixture of ice
and frozen eutectic solution.
When the temperature of a brine with a concentration above the eutectic falls below the freezing
point, salt crystallizes out and the concentration of
the residual solution decreases until at the eutectic
temperature the remaining solution reaches a eutectic concentration. Below this temperature the
mixture solidifies to form a mechanical mixture of
salt and frozen eutectic solution.
This chapter includes a discussion of these
brines, also tables and charts indicating properties.
I
FREEZlNG
POINT
A\\\\W
LlOUlD
SOLUTION
OF EUTECTIC
-
CONCENTRATION,PERCENT
SOLUTE
IN
MIXTURE
Courtesy of ASHRAE Guide and Data Book 1963
F IG . 11
BRINE
-
BR I N E M I X T U R E
SELECTION
The selection of a brine is based upon a consideration of the following factors:
1. Freezing Point - Brine must be suitable for
the lowest operating temperature.
.2. Application - When using an open piping
system, the possibility of product contamination by the brine should be checked.
3. Cost - The initial charge and quantity of
make-up required are factors in the determination of costs.
4. Safety - Toxicity and flammability of brine.
5. Thermal Performance - Viscosity, specific
gravity, specific heat and thermal conductivity
are utilized to determine thermal performance.
6. Suitability - Piping and system equipment
material require a stable and relatively corrosive-free brine.
7. Codes - Brine must be acceptable by codes,
ordinances, regulatory agencies and insuror.
4-24
l’.\IC’l‘
I
.
l~1:I~I~I(;I:I~.\N’I’S,
I1I<INI:S.
0II.S
TABLE I-TYPICAL BRINE APPLICATIONS
Snow Melting
X
X
X
X
Low Temperature (Special)
lea Cream
The most common brines are aqueous solutions
^ calcium chloride or sodium chloride. Although
,,oth of these brines have the advantage of low cost,
they have the disadvantage of being corrosive. To
overcome corrosion, an inhibitor may be added to
the brine. Sodium dichromate is a satisfactory and
economical inhibitor. Sodium hydroxide is added
to keep the brine slightly alkaline.
Sodium chloride is cheaper than calcium chloride
brine; however, it cannot be used below its eutectic
point of -6 F. It is preferred where contact with
calcium chloride brine cannot be tolerated, for
example, with unsealed foodstuffs. The use of calcium chloride of commercial grade is not satisfactory below -40 F.
Systems using aqueous solutions of alcohol or
glycol are more susceptible to leakage than those
using salts. A disadvantage of alcohol is its flatimability. It is utilized mainly in industrial proc, =qses where similar hazards already exist, and in
e same temperature range as the salts (down to
-40 F). The toxicity of methanol water (wood
alcohol) is a disadvantage. Conversely, the nontoxicity of ethanol water (denatured grain alcohol)
is an advantage.
Corrosion inhibitors should b’e used with alcohol
type brines as recommended by the manufacturer
of the alcohol.
Aqueous solutions of glycol are utilized mainly
in commercial applications as opposed to industrial processes. Ethylene and propylene glycol possess
equal corrosiveness which an inhibitor c&n neutralize. Galvanized surfaces are particularly prone
to attack by the glycols and should be avoided.
An inhibitor and potable water are recommended
for making up glycol brines. The glycol manufacturer sho~~ltl be consulted for inhibitor recommen.
X
X
X
X
dations. Some manufacturers have a brine sample
analysis service to assist in maintaining a satisfactory
brine condition in the system. Heat, transfer glycols
are available with nonoily inhibitors which do not
penalize heat transfer qualities (sodium nitrite or
borox).
Glycols can be used as heat transfer media at
relatively high temperatures. With stabilizers, glycol
oxidization in air at high temperatures is eliminated
for all practical purposes.
Ethylene glycol is more toxic than propyle;e glycol, but less toxic than methanol water. Propylene
glycol is pieferred to ethylene glycol in food freezing
for example.
Chlorinated and fluorinated hydrocarbons are
expensive and are used in very low temperature
work (below -40 F) .
Table 8 presents typical applications for the vario u s brives.
Load, brine quantity and temperature rise are
all related to each other so that, when any two are
known, the third may he found by the formula:
Load (tons) =
g-pm
X
temp
rise (F) X
sp gr
X
Cp
24
where:
sp gr = specific gravity of I)rine
Cp = specific heat of brine (fitu,/ll)-F)
PIPING
All materials in the piping system including
flange gaskets, valve seats and packing, pump seals
and other specialities must be compatible with the
brine. Copper tubing (except for the salt brines)
and standard steel pipe are suitable for general use.
The pump rating and motor horsepower should
be based on the particular brine used and the actual
operating temperature.
:HAPTER
4-25
2. BRINES
FRICTION LOSS
To determine the friction loss in a brine piping
system, the engineer should first calculate the loss
as if water were being used. A multiplier is then
used to convert the calculated loss to the actual
loss for the brine system. The multiplier is calculated as follows:
Refer to Chart I.
For a Reynolds number of 9.52 X 1Oa and a relative roughness of .000104, the chart indicates friction factor fb = .031.
Specilic gravity of fresh
55 F = 1.00
water at a mean temperature of
Refer to Chart 28.
Viscosity of fresh water at a mean temperature of 55 F = 1.2
centipoises.
7740 x ,575 x 4 . 2 9 x 1.00
Multiplier = sp gr X 7
w
where:
sp gr = specific gravity of brine
fb = friction I actor for the brine
f, = friction factor for water at the brine
Re (water) =
velocity
fb
Friction multiplier = sp gr (brine) X r
,031
W
= 1.05 x .o’L7 = 1.21
Brine friction loss = 1.21 X 7.5 psi = 9.08 psi or
7540 X d X v X sp gr
=
P’
where:
9.08 x 2.31
1.05
= 2 0 . 0 ft brine
PUMP BRAKE HORSEPOWER
d = inside pipe diameter (in.)
v = brine velocity (ftjsec)
To determine the horsepower required by a pump
with brine, the following formula may be used:
lb/cu ft
sp gr = specific gravity of brine =62.5
p’ = viscosity (centipoises) =
= 1 5 , 9 0 0 = 1.59 x 10’
Refer again to Chart 1.
For a Reynolds number of 1.59 X 10’ and a relative roughness of .000104, the chart indicates a friction factor f, = ,027.
Friction factor is determined from the Reynolds
number. The Reynolds number is:
Re =
1.2
ahsolute viscosity, lb/(hr)
gpm X total head (ft brine) X sp gr
(ft)
bhp =
2.42
3960 X eff
where:
Example 1 illustrates the use of the multiplier to
determine the brine friction loss thru a heat transfer
coil.
am
= gallons/min. of brine
total head = total pump head (ft brine)
= specific gravity of brine
sp gr
eff
= pump efficiency
Example 7 - Friction Loss Multiplier
Given:
A 5/8 in. copper tube coil with a circuit water velocity of
4.29 ft/sec and a pressure drop of 7.5 psi.
Mean water temperature = 55 F.
Find:
Fric’ ‘II loss multiplier and pressure drop when using
et],.
ie glycol at a mean brine temperature of 92.5 F and
417; sblution by weight at the same circuit liquid velocity.
Solution:
Refer to Chart 1.
E
.00006
T = 7ji5 = .000104
where:
f = absolute roughness of drawn tubing
d = inside diameter of sh in. copper tubing
Refer to Chart 19.
Specific gravity of ethylene glycol at a mean brine temperature of 92.5 F and 41% solution by weight is 1.05.
Refer to C h a r t 18.
Viscosity of ethylene
2.1 centipoises.
7740 x ,575
Re=
glycol
at
the
same
conditions
x 4 . 2 9 x 1.05
2.1
= 9520 = 9 . 5 2 x 108
equals
BRINE PROPERTIES
Specific gravity, viscosity, conductivity, specific
heat, concentration, and freezing and boiling points
are important factors in the consideration of liquids
other than water suitable for cooling and heating
purposes. High values of specific gravity, conductivity aixd specific heat, and low values of viscosity,
promote a high rate of heat transfer. High values
of specific gravity and viscosity result in high pumping head and consequentIy high pumping costs.
High specific heats are desirable in that they reduce the quantity of liquid required to be circulated or stored for a given duty. Low viscosities
are desirable from a standpoint of both rate of
heat transfer and low pumping costs. They are
particularly desirable at the lower temperatures
where the viscosity increases.
Table 9 is a tabulation of the various brines
covered in this chapter, giving the properties of
these brines at different temperatures and suitable
concentrations. Charts 2 to 28 present the viscosity,
specific gravity, specific heat and thermal conduc-
TABLE 9-BRINE PROPERTIES
olutior I
bv WI)
I )entity
(%I
12
12
15
20
25
30
(1 b/cu ft)
68.2
30
Sodium Chloride
Calcium Chloride
Methonot Water
Ethanol Water
Ethylene Glycol
Propylene Glycol
69.2
61.5
61.0
64.7
64.5
.a3
I .oo
1.04
.92
.94
21
20
22
25
35
40
72.0
74.8
15
Sodium Chloride
Calcium
Chloride
Methanol Water
Ethanol Water
Ethylene Glycol
Propylene Glycol
30
.72
.97
1.02
- 3 0
Calcium Chloride
Methanol Water
Ethanol Water
Ethylene Glycol
Propylene Glycol
25
35
36
45
50
30
45
52
55
60
78.4
- 5
Calcium Chloride
Methanol Water
Ethanol Water
Ethylene Glycol
Propylene Glycol
Temp.
(F)
Brine
60.4
61.0
66.0
65.3
60.0
60.6
67.4
66.5
82. I
60.0
59.5
69.0
67.2
Specific
Heat
Btu/lb-F
.8b
.8b
.09
.b7
.89
.95
.79
.a3
.b3
30
.a1
.73
.77
Thermal
Cond.
(Btu/hr
rq ft-F/ft)
.28
.32
.28
.27
.30
.26
.25
.31
.26
.25
.28
.24
‘29
.23
.22
.25
.23
-28
.22
.19
.22
.21
=coefficient
of heat transfer between brine and surface (Btu/hr-rq
ft-F),
(ft/sec),
at Re= 3500 for ,554 in. ID tubing.
t Vb = minimum brine velocity
IAbove
IO ftjsec.
I;
hb
tivity of the brines for various mean brine temperntures and compositions.
Note that specific gravity for propylene glycoI
(Chart 23) in the composition range of 507” t o
ircorlty
,eering
Point
,oiling
Point
2.2
2.4
3.2
5.5
3.7
8.0
W)
17.5
19.0
13.5
12.0
12.9
13.0
(F)
215.
213.
187.
189.
217.
216.
4.2
4.8
5.3
8.2
6.8
20.0
I .o
1.0
4.5
4.5
0.0
- 4.2
10.3
9.9
13.5
17.2
80.0
[centi-
hb
*
vb
t
Relative
Cart per
Gal. of
Solution
2.55
2.62
2.45
2.37
2.52
2.47
Tii971
781
621
775
525
1.61
1.78
2.63
4 60
2.92
6.35
1
3
13
20
42
43
216.
214.
182.
187.
2 I 9.
218.
2.57
2.77
693
730
599
504
576
103
2 . 9 0
3.28
4.44
6.05
5.25
1
5
19
25
60
58
-21.0
- 22.0
- 16.0
- 15.5
- 29.0
215.
176.
183.
223.
222.
2.85
2.82
2.62
2.82
2.72
-
216.
2.90
3.13
3.1 1
2.98
2.90
soiser)
27.8
18.0
20.2
75.0
700.0
at
7
T0 ipm/ton
47.0
45.0
50.0
43.0
55.0
velocity
I 110
171.
179.
227.
227.
r .554 iin. I D
deg
rise
2.56
2.41
2.65
2.58
513
98
97
103
98
110
91
83
93
91
-L
6.75
8.40
A
6
30
35
78
75
8
39
50
97
90
tubing.
IOOY” (same mean brine temperature, F) is the
same for two compositions. Specific gravity alone,
therefore, is not a reliable method of determining
the solution composition of this brine.
.
.
-..
.09
.OB
i
iiiii-liiii
I/III II I I Ill1
i I I //iii IliCi II /I/~-H~--I-.i~~I~~~I---l~l
02
,015
.Ol
.OOB
!!!!!!!>!I
!!
I
L
I
!!!‘-+““-i’..’
III1
TrtT
I
/
006
l-m
,004
,001
0008
0006
IIIIIIII I
I
I
NWII
I IIIII
! I, I,, II,
,
106
2
3
4 5 6
8 IO’
2
3
4 5 6 8
\
‘6
7740(d) Y (sg)
2
=.000.005
:
i
.000,001
=
P’
4
REYNOLDS
NUMBER (f?t?)
56
8
d = inside pipe diameter
Surface
c (in.)
Drawn tubing (very smooth surfaces of all kinds)
0.00006
sg
= specific gravity
Commercial steel or wrought iron
0.0018
/.L’
= viscosity (centipoises)
Galvanized
0.006
iron
v = brine velocity (ft/sec)
E = absolute roughness
I(
WIII
z
w
$
”
:
n
CHART 2-SODIUM CHLORIDE-VISCOSITY
20%
-10
0
IO
MEAN
SOLUTION BY
20
BRINE
TEMPERATURE
30
(F)
40
50
CHART J-SODIUM CHLORIDE-SPECIFIC GRAVITY
I .I8
I. 16
0
*With reference
to
F water.
20
MEAN
BRINE
40
TEMPERATURE
60
(F)
-I--30
I’\I<‘l‘
I.
I~l~:I~‘I~l~~1~:l1.\N’I’S.
l~l<I~lI:.S.
OILS
CHART 4-SODIUM CHLORIDE-SPECIFIC HEAT
.
%
.
I
j
-1..
-4
,I /
/
,..I
.
.
.,...
,
.9l
.
.9c
.8C
.8E I-
j-
. . . . i
,\6-..
.
i
.
.
i.
.{
I
.81
!O
0
+ 20
MEAN,
BRINE
40
60
TEMPERATURE (F)
80
Courtesy
of Dow Chcmid
Co.
CHART 5--SODIUM CHLORIDE-THERMAL CONDUCTIVITY
z
cu.
0
m
I
a
I
\
c
?
m
/
.30 ,
e
.25
.
MEAN BRINE TEMPERATURE(F)
30
50
70
CHART 6-CALCIUM CHLORIDE-VISCOSITY
MEAN
B R I N E T E M P E R A T U R E (F)
30
20
cn
u
fn
0
a
t
2
w
u
>
I:
..
9-.
8
7
6
M E A N B R I N E T E M P E R A T U R E (F)
.
CHART 7-CALCIUM CHLORIDE-SPECIFIC GRAVITY
.
__
i--
- 4 0
- 20
0
MEAN
*With reference to 60 F water.
20
BRINE
4 0
60
80
T E M P E R A T U R E (F)
Courtesy of Dow
Chemiral
Co.
CHART S-CALCIUM CHLORIDE-SPECIFIC HEAT
-40
-20
0
MEAN BRINE
20
40
TEMPERATURE (F)
60
80
Courtesy of Dow Chemical Co.
CHART 9-CALCIUM CHLORIDE-THERMAL CONDUCTIVITY
.33
.
^;,.
.
------+.,--
.28
.27
-40
-30
-20
-10
0
IO
20
30
MEAN BRINE TEMPERATURE
4 0
( F)
5 0
6 0
7 0
80
CHART
IO-METHANOL
BRINE-VISCOSITY
70
60
50
I
- 50
- 4 0
- 3 0
- 2 0
- IO
MEAN
BRINE
0
IO
2 0
TEMPERATURE
Courtesy
of
30
4 0
50
(F)
Carbide
and
Carbon Clmnicals
Corporation
(:H,\1”I‘EIi
2.
4-37
11IIINES
-
-
-
-
CHART ll--METHANOL BRINE-SPECIFIC GRAVITY
.8
.8
- 5 0
- 4 0
-30
‘W ith reference to 60 F water.
- 2 0
MEAN
-10
BRINE
0
IO
TEMPERATURE
Courtesy
20
(F)
30
4 0
5 0
of Carbide and Carbon Chemicals Corporatim
l'.\I<'I‘
4-38
CHART
12--METHANOL
I. IIEl~lil(;l-li,\N~1‘S,
BRINE--SPECIFIC HEAT
1.10
1.05
I.OC
c
!
0.95
m
-I
'= 0.9(
im
" 0.8!
2
g 0.81
0
ic
E 0.7!
w
a
m
0.7f
0.6:
0.6(
0.5:
-60
-50
-40
-30
MEAN
-20
-IV
B R I N E TEMPER:T”RE’~F~
zv
I 3v
IIRINES.
OILS
CHART 13--METHANOL
BRINE-THERMAL CONDUCTIVITY
.32
.I0
- 5 0
- 4 0
- 3 0
- 2 0
MEAN
-10
BRINE
0
IO
20
T E M P E R A T U R E IF)
30
4 0
50
CHART 14-ETHANOL BRINE-VISCOSITY
100
90
80
70
60
v)
W
tn
3
c
iii
0
0
0-J
?I
0
a
2
c
z
W
0
‘I,’
- 50
- 4 0
- 3 0
0
IO
-10
- 20
M E A N B R I N E T E M P E R A T U R E (F)
20
30
“““a
4 0
.
CM,\I”I‘I~K
_I
. ‘).
IiKIlNLS
4-41
---_
CHART 15-ETHANOL BRINE-SPECIFIC GRAVITY
.
j
1
.
I
JPOINT
i
!
.
.J
_
,
?‘”
:
.._...
;:;;. J
_:
.95
u .90
.61 ,
._
.6C ) -
-60
-20
MEAN
*With
reference
to GO F water.
0
BRINE
20
TEMPERATURE
40
60
(F)
Extrap&ted
values from International Critical Tables
CHART 16-ETHANOLBRINE-SPECIFIC HEAT
m
i
u
CHART 17-ETHANOL BRINE-THERMAL CONDUCTIVITY
4-44
l’.\l<‘I‘
,I. I~l:l~Kl(;I:I1,\N’I‘S.
IIKINES, O I L S
\
CHART 1 II-ETHYLENE GLYCOL-VISCOSITY
3000
2000
FREEZING
/
1000
POINT
CURVE
800
600
400
300
100
8 0
6 0
u-l
w
v)
4 0
6
a
3 0
F
z
w
2 0
0
81”
I
6
I
Q
CURVE
0.8
0.6
0 .4
0.3
0.2
0. I
-100
- 50
0
MEAN
100
50
BRINE
TEMPERATURE
150
200
(F)
From Glycols, Properties and Uses, Dow Chemical Co. 1961
CHART 19-ETHYLENEGLYCOL-SPECIFIC GRAVITY
1.10
1.04
1.02
1.00
.99
MEAN
*With reference to GO F water.
BRINE
TEMPERATURE
(F)
CHART 20--ETHYLENE
GLYCOL-SPECIFIC HEAT
CHART 2 I-ETHYLENE GLYCOL-THERMAL
CONDUCTIVITY
.4c
I-.
._
, -
I -
Y
._
5 ,
II
.
.If 5 _
.IC
- 50
0
50
MEAN
100
BRINE
150
TEMPERATURE
200
250
(F)
From Glgcol~,
Union Carbide Chemicals Co. 1958
4-48
l'.\ll'I‘
1. 1~1:I~l~I(~1~II,\N'I',S. lII<lNli,S, OILS
CHART 22-PROPYLENE GLYCOL-VISCOSITY
600
400
300
200
.
100
80
-
60
IO
8
6
I
0.6
0.6
0.4
0.3
0.2
0.1
0
50
100
150
20
MEAN BRINE TEMPERATURE (F)
From Gl~cols,
ProWlies nttd User, Dow Chenlical
Co. I!)(il
CHART 23-PROPYLENE GLYCOL-SPECIFIC GRAVITY
1.06
2c.0
I
1.02
I
FREE21
- POINT
CURVE
3.99
I.01 .
1.00 -
0.99
-I
0.90
-40
-20
0
With reference to GO F water.
20
40
60
80
100
120
M E A N B R I N E T E M P E R A T U R E (F)
From Glycols, ProperGes
140
160
180
and U s e s , DO W Chemical Co. 1961
4-750
l’.\l<‘l’ ,I. li~:I~l~i~;I:I~.\N1‘S. I(I<INL:.S, O I L S
CHART 24-PROPYLENE GLYCOL-SPECIFIC HEAT
1.0
0.9
.
0.6
- 5 0
0
5 0
MEAN
BRINE
100
TEMPERATURE
150
(F)
From Clycols,
2 0 0
2 5 0
Union Carbide Ctwmicnls Co. 1958
CHART 2S-PROPYLENEGLYCOL-THERMAL CONDUCTiVlTY
3
432
I’/\R’I
I.
I~EI~I~IC;EK.\N’I‘.S,
IIKINES,
CHART 26-TRICHLOROETHYLENE-PROPERTIES
y .07
z
5
t-v
2
5 2.2
0
4
I 2.1
a
w
= 2.c
1.9
I .E
1.60
1.7
;
w
07
I.6
E
1.5
F
z
z
1.4
c
;;
I.;!
0
0
z
1.2
I.1
I.C
0.9
- IZU
-100
- 00
- 6 0
MEAN
*With
reference t o G O F water.
BRINE
- 4 0
TEMPERATURE
-20
(F)
OILS
CHART 27-REFRIGERANT ll-PROPERTIES
-100 -90
-00
-70
-60
-50
MEAN
*With reference to GO F water.
-40
-30
-20
BRINE
TEMP
(F)
-10
0
IO
20
30
40
CHART 28-WATER-VISCOSITY
1 . 9
I.0
.
. . . .I
.
. . . . j . .
Specific heat = 0.940 Btu/lbF
Specific gravity = 1.025
‘l7vzrmal conductivity -See Chart 5. IJsc 3’;6 solution.
1.7
1.6
.
30
40
50
60
70
00
90
MEAN BRINE TEMPERATURE
100
(F)
110
120
CHAPTER 3. REFRIGERATION OILS
I liis chapter covers general classifications and
quality of lubricating oils that are important in
refrigeration. The recommendation of oils to be
used in a refrigeration system is primarily the responsibility of the refrigeration system manufacturer. However, it is important for an engineer to
understand the basis of the selection of these oils
in order to properly apply them in the field.
CLASSIFICATION
Oils classified by source fall in three main groups:
animal, vegetable and mineral. Animal and vegetable oils are called fixed oils because they cannot
be refined without decomposing. They are unstable
and tend to form acids and gums that make them
unsuitable for refrigeration purposes.
There are three major classifications of mineral
o;‘ laphthene base, paraffin base, and mixed base.
W . ..n distilled, a naphthene base oil yields a residue
of heavy pitch or asphalt. California oils, some Gulf
Coast and heavy Mexican oils are in this class. A
paraffin base oil yields a paraffin wax when distilled.
The best sources of paraffin base oils are Pennsylvania, Northern Louisiana, and parts of Oklahoma
and Kansas. The mixed oils contain both naphthene
and paraffin bases. Illinois and some mid-continent
oils are in this class.
Experience has shown that the naphthene base
oils are more suited for refrigeration work for three
main reasons:
1. They flow better at low temperatures.
2. Carbon deposits from these oils are of a soft
nature and can easily be removed.
3. They deposit less wax at low temperatures.
When obtained from selected crudes and properly
refined and treated, all three classes of mineral oil
can be considered satisfactory for refrigeration use.
PROPERTIES
To meet the requirements of a refrigeration system, a good refrigeration oil should:
1. Maintain sufficient body to lubricate at high
temperature and yet be fluid enough to flow
at low temperature.
2. Have a pour point low enough to allow ROW
at any point in the system.
3. Leave no carbon deposits when in contact
with hot surfaces encountered in the system
during normal operation.
4. Deposit no wax when exposed to the lowest
temperatures normally encountered in the
system.
5. Contain little or no corrosive acid.
6. Have a high resistance to the flow of electricity.
7. Have a high flash and fire point to indicate
proper blending.
8. Be stable in the presence of oxygen.
9. Contain no sulfur compounds.
10. Contain no moisture.
11. Be light in color, to indicate proper refining.
As lubricating oils for refrigeration compressors
are a specialty product, they require consideration
apart from normal lubricants. The emphasis in this
chapter is on oil used in refrigeration. Do not consider the emphasis as applicable to lubricants in
general.
l’.\K’I‘
4-56
FIG.
12
-
L UBRICATION CHARACTERISTICS
The characteristics of oil for refrigeration (not
,ecessarily
in order of importance) are these:
1. Viscosity
2. Pour point
3. Carbonization
4. Floe point
5. Neutralization
6. Dielectric strength
7. Flash point
8. Fire point
9. Oxidation stability
10. Corrosion tendency
11. Moisture content
12. Color
VISCOSITY
Viscosity or coefficient of internal friction is that
qroperty of a liquid responsible for resistance to
low; it indicates how thick or thin an oil is.
The purpose of an oil is to lubricate bearing or
rubbing surfaces. If the oil is too thin, it does not
stay between the rubbing surfaces but is forced out,
leaving no protective film. If the oil is too thick, it
causes drag and loss of power, and may not be
able to flow between the bearing or rubbing surfaces.
Friction loss f is illustrated in I;i,n. 12 as a function
of viscosity Z, speed N in revolutions per unit time,
and load P per unit area.
Viscosity is usually measured in terms of Saybolt
Seconds Universal (SSU). Under standard temperature conditions, oil is allowed to flow thru a carefully calibrated orifice until a standard volume has
passed. The number of seconds necessary for the
given volume of oil to How thru the orifice is the
viscosity of the oil in Saybolt Seconds Universal.
.
FIG.
13
I . I~I~~I~I~IC;EII.\N~l‘S,
- EFFECT 01: TEMPIIRAIIJRE
I\I<INES,
ON
OILS
V ISCOSITY
The higher the viscosity, the more seconds it takes
to pass thru the hole, or the higher the viscosity,
the thicker the oil.
Viscosity is affected by temperature (Fig. 13), thus
making it an important characteristic of refrigeration oils. The viscosity increases as the temperature
decreases, or the lower the temperature the thicker
the oil. In low temperature applications, this thickening of oil with its increasing resistance to How is
a major problem. As low temperatures occu; in the
evapoiator, an oil that is too viscous thickens and
may stay in the evaporator, thus decreasing the heat
transfer and possibly creating a serious lack of
lubrication in the compressor.
Oil may thin out or become less viscous at high
temperatures. Too warm a crankcase may conceivably thin the oil to a point whe:e it can no longer
lubricate properly. A refrigeration oil must maintain sufficient body to lubricate at high temperatures and yet to be Huid enough to How at low
temperatures. Oil should be selected which has the
lowest viscosity possible to do the assigned job.
Viscosity is also affected by the miscibility of the
oil and the refrigerant. The miscibility of oil and
refrigerants varies from almost no mixing (with Refrigerant 717, ammonia) to complete mixing (with
some halogenated hydrocarbons such as Refrigerant 12).
Refrigerant 717 has almost no effect on the viscosity of a properly refined refrigeration oil. As it is
not miscible, there is no dilution of oil and therefore no change in viscosity.
In the case of miscible refrigerants such as Refrigerant 12, the refrigerant mixes with and dilutes
the oil, and lubrication must be performed by this
mixture. This mixing reduces the viscosity of the oil.
.
CHART
29-OIL-REFRIGERANT
VISCOSITY
- 2 0 0 0
4’
z
500
T
5
z
200
IO0
z
s
In
r
6 0
d
m
2
2
z
3
40
FIG. 14 - POUR I'OINT TUIE
35
8
0)
5
-20
0
20
40
60
60
100
TEMPERATURE
120
140
160
160
210
(Ft
P E R C E N T R E F R I G E R A N T 12 I N O I L
Chart 23 shows viscosity change for mixtures of
oil and Refrigerant 12. For example, oil at 40 F containing 20y0 Refrigerant 12 by weight has a viscosity
of approximately 150 SSU.
When the amount of Refrigerant 12 is increased
to 40%, the SSU is reduced to 45.
When oil and refrigerants are miscible, the oil is
carried thru the system by the refrigerant. It is
imperative that the oil be returned to the compressor. Keeping low side gas velocity up assures this
proper oil return. With a completely miscible refrigerant, the oil is diluted sufficiently even at low
temperatures to keep the viscosity iow and to allow
the oil to return easily with the refrigerant to the
compressor.
:re is a third group of refrigerants whose
miscibility with oil varies. For example, Refrigerant
22 is completely miscible with oil at high temperatures, but at low temperatures it separates into two
layers with the oil on top. When designing or selecting equipment for this group of refrigerants, great
care must be taken to allow for low temperature
separation of oil and refrigerant.
Oil separation in a flooded cooler necessitates the
use of an oil bleed line from the bottom of the
cooler to the suction loop. Because the oil is lighter
than Refrigerant 22 and floats on top of the liquid
refrigerant, an auxiliary oil bleed line from the side
of the cooler is required.
POUR POINT
The pour point of an oil is that temperature at
which it ceases to flow.
FIG. 15 - POURPOINT
Pour point is simple to determine. Using the
apparatus shown in the Fig. 14, the selected batch
of oil is slowly cooled under test conditions until
the oil no longer flows. This temperature is the
pour point.
Figure 15 shows two oils with different pour
points which have been cooled to the same temperature (-20 F). On the left the oil with a -40 F pour
point flows freely. On the right the oil with the 0 F
pour point does not flow.
Pour point depends on the wax content and/or
viscosity.
With all refrigerants some oil is passed to the
evaporator. Regardless of how small an amount,
this oil must be returned to the compressor. In order
that it may be returned, it must be able to flow
thruout the system.
Oil pour point is very important with nonmiscible and partly miscible refrigerants; with the
miscible refrigerants the viscoscity
of the oil refrigerant mixture assumes greater importance as
shown in Chart 29, Oil-Refrigeuznt Viscosity.
COPPER
SLUDGE
FIG. 1G -
&WON
PLATING
Del~osu
CARBONIZATION
111 refrigeration oils can be clecon~posed
by heat.
When such action takes place, a carbon deposit
remains.
Carbonization properties of an oil are measured
by the Conradson Carbon Value. This value is found
by heating and decomposing an oil until only the
carbon deposit remains. The ratio of the weight of
the carbon deposit to the weight of the original oil
sample is the Conradson Carbon Value.
Hot surfaces within the refrigeration system sometimes decompose the oil. The carbon remaining is
hard and adhesive in paraffin base oils, and forms
sludge.
Naphthene base oils form a light fluffy carbon
which, though a contaminant, is not as damaging
as the hard carbon. However, neither type of carbon
deposit is desirable as there is some indication that
a relationship exists between oil breakdown, carnization and copper plating (Fig. 16).
A good oil should not carbonize when in contact
with hot surfaces encountered in the system during
normal operation. A refrigeration oil should have as
low a Conradson Carbon Value as practical.
FIG. 17 - FLOC TEST
ture at which these clusters are first noticeable to the
unaided eye is the floe point.
The free wax that is formed when a refrigeration
oil is cooled can clog metering devices and restrict
flow.
Wax normally deposits out in the colder parts of
the system such as in the evaporator and its metering device (Fig. IS). Wax in the evaporator causes
some loss of heat transfer; wax in the metering device can cause restriction or sticking.
A good refrigeration oil should not deposit wax
when exposed to the lowest temperatures normally
encountered in the refrigeration system.
NEUTRALIZATION
Almost all refrigeration oils have some acid tendencies. Nearly all oil contains material of uncertain
composition referred to as organic acids. These are
FLOC POINT
All refrigeration oils contain some wax though
the amount varies considerably. As the temperature
of the oil decreases, the solubility of the wax also decreases. When there is more wax present than the oil
can hold, some separates
and precipitates.
The method used to determine the waxing
tendencies of a refrigeration oil is the Aoc test
17). A mixture of 10% oil and 90% Refrigerant 12
in a clear container is cooled until the wax starts to
separate, turning the mixture cloudy. As cooling
continues, small clusters of wax form. The tempera-
FIG. 18 - WAX DEPOSITS
.
FIG. 19 - DIELECTRIC TESTING
usually harniless anti si~oultl not be conl‘used with
mineral acids which are harmful.
T!- - neutralization number is a measure of the
ama
_ of mineral acid, and is determined by
measuring the amount of test Iluid that must be
added to the oil to bring it to a neutral condition.
A low neutralization number means that few acids
are contained in the oil.
Improper refining may leave a large proportion
of corrosive acid present in an oil. A low neutralization number indicates a low acid content. Acids may
corrode interior parts of the system; they react with
motor insulation and other materials to form sludge
which can eventually cause a complete system breakdown. A low neutralization number is highly desirable in refrigeration oils.
DIELECTRIC
STRENGTH
Dielectric strength is a measure of the resistance
of an oil to the passage of electric current. It is
measured in kilovolts on a test cell as shown in
Fig. ‘?. The poles in the cell are a predetermined
dist, -e apart. They are immersed in the oil so that
current must pass thru the oil to flow from one pole
to the other. The kilovolts necessary to cause a spark
to jump this gap is known as the dielectric rating.
Good refrigeration oils normally have a rating of
over 2.5 kilovolts.
A dielectric rating is important because it is a
measure of impurities in the oil. If the oil is free
of foreign matter, it has a high resistance to current flow. If the oil contains impurities, its resistance
to current flow is low.
The presence of foreign matter in a refrigeration
system is sufficient reason for considering this test
valuable. Hermetic motors make a high kilovolt
rating necessary for refrigeration oils since a low
kilovolt rating may be a contributing factor to
shorted windings.
FIG. 20 - FLASH
AND
FIRE POINT TESTS
FLASH POINT AND FIRE POINT
The flash point of an oil is that temperature at
which oil vapor flashes when exposed to a flame.
The fire point is that temperature at which it continues to burn.
The apparatus shown in Fig. 20 heats the oil
while a small gas flame is passed closely over the
surface of the oil. When a flash of fire is noted at
some point on the surface, the flash point has been
reached. The apparatus continues to heat the oil
until it ignites and continues to burn. This is the
fire point.
The flash point of a good refrigeration oil is well
over 300 F. Temperatures obtained in the normal
refrigeration system rarely reach this point. The
test for flash and fire points is important as it is
a means of detecting inferior blends.
It is possible to get an acceptable viscosity reading for a refrigerant oil by mixing a small amount
of high viscosity oil with a large amount of low
viscosity oil. The viscosity of the mixture indicates
a satisfactory oil when actually the low viscosity
oil is inferior and breaks down under normal use.
Fortunately, this can be detected using the flash
and fire point test which indicates the inferiority of
the low viscosity oil.
OXIDATION
STABILITY
Oxidation stability is the ability of refrigeration
oil to remain stable in the presence of oxygen. The
Sligh Oxidation Test is used to determine this stability (Fig. 21).
While exposed to oxygen in the flask, the oil is
heated to a high temperature for an extended period
of time. The solid sludge that is formecl in the flask
F1c.21
-SLICH
OXIDATION TEST
FIG. 23 -
TEST
CORROSlON
.
SLUDGE
FREEZE-UP
Flc.22 -OILBREAKDOWN
is weighed and reported as the Sligh Oxidation
Number.
When air enters a system, some moisture generally accompanies it. The combination of moisture,
air, refrigeration oil and discharge temperatures
c educes acid which creates sludge (Fig. X). If oil
<
‘nas a low Sligh Oxidation Number, the oil breakdown to acid and sludge is quite slow.
CORROSION TENDENCY
The corrosion tendency of a refrigeration oil is
measured by the copper strip corrosion test (I;ig.
23). This test is intended to indicate the presence
of undesirable sulfur compounds in the refrigeration oil.
A strip of polished copper is immersed in an oil
sample in a test tube. This is subjected to temperatures around 200 F. After about 3 hours, the copper
is removed from the oil, cleaned with a solvent,
and examined for discoloration. If the copper is
tarnished or pitted, then sulfur is present in the oil.
Well refined oils rarely cause more than a slight
tarnishing of copper in this test.
ACIDS
F1c.24 - MOISTUREANDXIRINTHE
SYSTEM
A good refrigeration oil should score negative in
the copper strip corrosion test. If it does not, it
contains sulfur in a corrosive form.
Sulfur alone is a deadly enemy of the refrigeration system but, in the presence of moisture, surfurous acid is formed, one of the most corrosive
compounds in existence. Though the sulfurous acid
converts immediately to sludge; this sludge is certain
to create serious mechanical problems.
MOISTURE CONTENT
Moisture in any form is an enemy of the refrigeration system; moisture contributes to copper plating,
formation of sludges and acids, and can cause freezeup (Fig. 24). No refrigeration oil should contain
enough moisture to affect the refrigeration system.
(;tI,\l’-I‘EK 3.
KEi~KIGEK,\-I‘lON
OILS
4-61
oil. These are believed to be the constituents in oil
that act as a solvent for copper. Therefore, the aim
is to refine the oil sufficiently to remove these hydrocarbons but not so much as to destroy the lubricating quality.
SPECIFICATIONS
FIG. 25 - COLOK?'ESTING
co1
The color of a refrigeration oil is expressed by a
numerical value that is based on comparison of the
oil with certain color standards. This is done with
the calorimeter shown in I;ig. 25.
The color of a good refrigeration oil should be
light but not water white. Continual refining of a
lubricating oil results in a watek white color. It also
results in poor lubricating qualities.
Under-refining leaves a high content of unsaturated hydrocarbons which darken and discolor an
1. For open and hermetic reciprocating compressors at standard air conditioning levels,
the following oil characteristics are typical:
Viscosity, 150 & 10 SW at 100 F
40 to 45 SSU at 210 F
Dielectric (min.), 25 kv
Pour Point (max.), -35 F
Flash Point (min.), 330 F
Neutralization Number (max.), .05
Floe Point (max.), - 70 F
2. For centrifugal compressors used for water
cooling at air conditioning levels, the following
are typical properties:
Viscosity, 300 2 25 SSU at 100 F
50 to 55 SSU at 210 F
Dielectric (min.), 25 kv
Pour Point (max.), 20 F
Flash Point (min.), 400 F
Neutralization Number (max.), 0.1
3. For special applications, consult the equipment
manufacturer.
P
CONTENTS
Preface.. . , . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
.
Part 1. LOAD ESTIMATING . . . . . . . . . . . . . . . . . . . .
1.
2.
3.
4.
5.
6.
7.
8.
.
.
.
.
.
.
. .
Part 2. AIR DISTRIBUTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
.......................
.......................
. .
3-l
.
.
. .
. . . .
3-l
3-19
3-43
3-81
.
. . . .
4-1
Part 3. PIPING DESIGN’:. . . . . . . . . . . . . . .
1.
2.
3.
4.
Piping Design-General . . . . .
Water Piping . . . . . . . . . . . . . .
Refrigerant Piping . . . . . . . . .
Steam Piping . . . . . . . . . . . . . .
. .
.
.
Part 4. REFRIGERANTS, BRINES, OILS
1. Refrigerants ..,.................................................
2. Brines . . . . . . . . . . . . . . . . . . . .
...........................
3. Refrigeration Oils . . . . . . . . .
.
. ...........................
.,
,._,
-
-.
2-1
2-1
2-17
2-65
.......................
1. Air Handling Apparatus .
2. Air Duct Design . . . . . . .
3. Roam Air Distribution . . . . .
I-1
1-1
1-9
l-25
1-41
1-59
l-89
l-99
l-l 15
.
Building Survey and Load Estimate . . . .
Design Conditions . . . . . . . . . . . . . , .
Heat Storage, Diversity and Stratification . .
Solar Heat Gain thru Glass , . . . . . . , . . . .
Heat and Water Vapor Flow thru Structures
Infiltration and Ventilation . . . . . . . .
Internal and System Heat Gain . , . . .
Applied Psychrometrics
. . . . . . , .
v
4-1
4-23
4-55
.-.
.-.-‘- ---- ~~~o~DITIoNINs-~~~~~.~.
1. Water Conditioning-General
2. Scale and Deposit Control .
3. Corrosion Control . . . . . . . .
4. Slime and Algae Control .
5. Water Conditioning Systems
6. Definitions . . . . . . . . . . . . . . .
.................
. . . . . . . . .
. .
c
. .
. .
....
....
....
....
....
5-1
5-l
5-11
5-19
5-27
5-31
5-47
’
I
Part
6
.
I . Pans
2. Air
3.
HANDLING
EQUIPMENT
.
:\c:ccssory
Apparatlls
6-1
Eqtlipmcnt
........................
Part 8. AUXILIARY EQUIPMENT
.
.
SYSTEMS
AND
.
.
.
.
7-l
.
. . . .
7-l
7-?1
7 - 3 3
.
. . . 7-47
. . . 7 - 5 5
. . . . . . . . . .
. . . . . .. . . .
.
.
8-l
. . . . . . . . . . . . .
Centrifugal Pumps . . . . . . . . . . . . . . . . . . . . . . . . . . .
Motors and Motor Controls . . . . . . . . . . . . . . . . . . .
Boilers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
Miscellaneous Drives . . . . . . . . . . . . . . . . . . . . . . . . . .
9.
G - 1 7
G - 4 5
. G-51
.
..........................
Eq~~iprncnt
1. Reciprocating Rcfrigcrntion Machine . . . . . . . . . . .
2 . Ccntrifugnl Rcfrigcration Machine . . . . . . . . . . . . .
’. 3. Absorption Rcfrigcrntion Machine . . . . . . . . . . . . . .
.I-. Comhinntion Absorption-Centrifugal System . . . . .
5 . IIcat Rejection Equipment . . . . .