Fan and Air System
Applications Handbook
AMCA International
Publication 200
Air Systems
Publication 201
Fans and Systems
Publication 202
Troubleshooting
Publication 203
Field Performance
Measurements of
Fan Systems
Forward
Publication 200
Publication 201
Publication 202
Publication 203
Air Systems
Fans and Systems
Troubleshooting
Field Performance
Measurements of Fan
Systems
Air Systems is intended
to provide basic
information needed to
design effective and
energy efficient air
systems. Discussion
is limited to systems
where there is a clear
separation of the fan
inlet and outlet and does
not cover applications in
which fans are used only
to circulate air in an open
space.
Fans and Systems
is aimed primarily at
the designer of the air
moving system and
discusses the effect
on inlet and outlet
connections of the fan’s
performance. System
Effect Factors, which
must be included in the
basic design calculations,
are listed for various
configurations. AMCA
202 and AMCA 203 are
companion documents.
Troubleshooting is
intended to help identify
and correct problems
with the performance
and operation of the air
moving system after
installation. AMCA 201
and AMCA 203 are
companion documents.
• System Pressure
Losses
• Fan Testing and Rating
• System Checklist
• Acceptance Tests
• The Fan “Laws”
• Fan Manufacturer’s
Analysis
• Test Methods and
Instruments
• Master Troubleshooting
Appendices
• Precautions
• Fan Performance
Characteristics
• System Effect
• System Design
Tolerances
Field Performance
Measurements of
Fan Systems reviews
the various problems
of making field
measurements and
calculating the actual
performance of the fan
and system. AMCA
201 and AMCA 202 are
companion documents.
• Air Systems
• Fan and System
Interaction
• Limitations and
Expected Accuracies
• System Effect Factors
• Calculations
Review Committee
These members contributed to the final review of the publications contained in this
Fan Application Manual.
Tom Berger
Pace Company, Division of York International
Rick Bursh
Illinois Blower, Inc.
Patrick Chinoda
Hartzell Fan, Inc.
Narsaiah Dasa
TLT-Babcock, Inc.
Gerald P. Jolette
AMCA Staff
Robert L. Lanier
Phelps Fan Manufacturing Co., Inc.
Tung Nguyen
Sutton G. Page
Scott Phillips
Emerson Ventilation Products
Austin Air Balancing Corp.
The New York Blower Company
Neil H. Rutherford
Delhi Industries, Inc.
Jack E. Saunders
Barry Blower/Snyder General Corp.
Paul R. Saxon
AMCA International
Erling Schmidt
Novenco, Inc.
Mark Schultz
American Fan Company
William Smiley
The Trane Company
James L. Smith
Aerovent, Inc.
Charles R. Voss
Phelps Fan Manufacturing Co., Inc.
Robert H. Zaleski
Acme Engineering & Manufacturing Corp.
Disclaimer
This manual has been prepared by the Air Movement
and Control Association, Inc. The information
contained in this manual has been derived from many
sources and is believed to e accurate. Please note
that the recommendations contained herein do not
necessarily represent the only methods or procedures
appropriate for the situation discussed, but rather are
inteded to present consensus opinions and practices
of the air movement and control industry which may
be helpful, or of interest to those who design, test,
install, operate or maintain fan-duct systems. Thus,
AMCA disclaimes any and all warranties, expressed
or implied, regarding the accuracy of the information
contained in this maual and further disclaims any
liability for the use or misuse of this information. AMCA
does not guarantee, certify or assure the performance
of any fan-duct system designed, tested, installed,
operated or maintained on the basis of the information
provided in this manual.
Air Movement and Control Association International,
Inc. will consider and decide all written complaints
regarding its standards, certification programs, or
interpretations thereof. For information on procedures
for submitting and handling complaints, write to:
Air Movement and Control Association International
30 West University Drive
Arlington Heights, IL 60004-1893 U.S.A.
or
AMCA International, Incorporated
c/o Federation of Environmental Trade Associations
2 Waltham Court, Milley Lane, Hare Hatch
Reading, Berkshire
RG10 9TH United Kingdom
Table of Contents
Publication 200
Air Systems
1
Introduction
3
Symbols and Subscripts
4
Properties of Air
5
Airflow
13
The Flow System
34
System Design and Tolerances
40
Annex A — SI / I-P Conversion Table
41
Annex B — Standard Atmospheric Data Versus Altitude Charts
43
Annex C — Psychrometric Density Tables
47
Annex D — Friction Charts
49
Annex E — Air Density Correction Factor Charts
Publication 201
Fans and Systems
51
Introduction
51
Symbols and Subscripts
51
Fan Testing
54
Fan Ratings
63
Catalog Performance Tables
66
Air Systems
74
System Effect Factor (SEF)
79
Outlet System Effect Factors
88
Inlet System Effect Factors
99
Effects of Factory Supplied Accessories
102
Annex A. — SI / I-P Conversion Table (Informative)
103
Annex B. — Dual Fan Systems - Series and Parallel
106
Annex C. — Definitions and Terminology
113
Annex D. — Examples of the Convertibility of Energy from Velocity Pressure
to Static Pressure
120
Annex E. — References
49
Table of Contents (continued)
Publication 202
Troubleshooting
122
Introduction
122
Procedure for Troubleshooting
122
Safety Precautions
122
System Checklist
128
Fan Manufacturer’s Analysis
130
Conclusion
131
Annex A. Noise
134
Annex B. Insufficient Airflow
136
Annex C. Airflow High
137
Annex D. Static Pressure Wrong
139
Annex E — Power High
140
Annex F — Fan Does Not Operate
141
Annex G — Premature Failure
142
Annex H — Vibration
Publication 203
Field Performance Measurement of Fan Systems
145
Introduction
145
Scope
145
Types of Field Tests
146
Alternatives to Conducting Field Tests
146
System Effect Factors
146
Fan Performance
146
Referenced Planes
147
Symbols and Subscripts
147
Fan Flow Rate
152
Fan Static Pressure
156
Fan Power Input
158
Fan Speed
158
Densities
Table of Contents (continued)
159
Conversion Calculations
160
Test Preparation
161
Precautions
161
Typical Fan-System Installations
165
Annex A — Field Test Examples
241
Annex B — Pitot-Static Tubes
242
Annex C — Double Reverse Tube
243
Annex D — Pitot-Static Tube Holder
244
Annex E — Static Pressure Tap
245
Annex F — Pitot-Static Tube Connections
246
Annex G — Manometer Data
248
Annex H — Distribution of Traverse Points
250
Annex J — Instrumentation Characteristics
252
Annex K — Phase Current Method for Estimating the Power Output of Three
Phase Fan Motors
254
Annex L — Estimated Belt Drive Loss
256
Annex M — Density Determinations
260
Annex N — Density Charts and Tables
269
Annex P — Diffusion at Fan Outlets
270
Annex R — Diffusion at Fan Outlets
274
Annex S — Typical Format for Field Test Data
275
Annex T — Uncertainties Analysis
Air Systems
1. Introduction
An air system is any assembly of ducts, filters,
conditioning devices, dampers, louvers, fans, etc.,
the main purpose of which is to move air from one
place to another in a controlled fashion. Most air
systems draw air from one space and discharge it
into another.
Air systems are often required to operate
satisfactorily in a wide range of environmental
conditions. The conditions which will be encountered
must be considered in the design of the ducts, pipes,
etc., which will contain the airflow and constitute the
boundary of the system.
1.1 Air system components
A typical air system may contain one or more of the
following (see Figure 1):
a) System inlet
b) Distribution system
c) Fan
d) Control device
e) Conditioning device
f) System outlet
1.1.1 System inlet. An air system usually includes
devices such as louvers, filters, screens, guards,
grilles, etc., where the air enters the system. These
are used for safety reasons as well as to inhibit the
entry of rain, dust, and other unwanted matter. Their
appearance may be important as they are usually
visible on the exterior of a structure.
1.1.2 Distribution system. Most air systems are
made up of ducts specially designed and constructed
to convey air from the system inlet(s) to the system
outlet(s). In some cases, enclosed spaces in the
structure such as plenums above ceilings or holes in
walls may be used to confine and direct the flow.
1.1.3 Fan. Understanding the design and opera-tion
of air systems begins with an understanding of the
various types of fans, their performance
characteristics, and their applications.
200
A fan is required in order to produce the pressure
differential which results in the flow of air through a
system. The fan must be carefully selected to meet
the specified airflow and pressure for proper system
operation. Different fan designs produce different
pressure-volume and fan power relationships, which
are critical to air system operation. Refer to Figure
4.2, AMCA Publication 201-90.
1.1.4 Control devices. In many air systems it is
necessary to regulate and control the flow through
the system in response to some monitoring signal,
usually temperature or pressure. It may be also
necessary to regulate the flow in the individual
branches of the system. Control devices such as
dampers function by controlling the amount of airflow.
In some cases, the output of the fan can be varied by
other methods (variable speed motor, variable inlet
vanes, variable pitch impeller, etc.)
1.1.5 Conditioning device. Most air systems are
designed to take air from the inlet and change its
condition before discharging it at the outlet. Changes
may include the temperature, humidity, pressure,
contaminant level and cleanliness, etc., of the air.
Many conditioning devices require outside energy
sources, for example, heating and cooling coils; other
components such as filters are passive devices and
have no external energy connection. All conditioning
devices increase the pressure drop across the
system and this effect must be considered in the
selection of the fan.
1.1.6 System outlet. An air system usually includes
a special component at the termination of the system
or at the end of each of the system's branches, such
as a simple screen or louver. In many cases the
distribution of the air at the outlet to the receiving
space is very important, e.g., in an occupied air
conditioned room. These systems require carefully
selected outlets and diffusing devices to achieve
desirable air motion and temperature conditions in
the conditioned space. Typical devices are ceiling
diffusers and grilles. In some cases these may
incorporate control devices such as dampers and
mixing boxes.
FAN
MAIN DISTRIBUTION SYSTEM (DUCT)
SYSTEM
INLET
BRANCH
DUCT
COIL
FILTER
LOUVER
DAMPER
DIFFUSER
SYSTEM
OUTLET
Figure 1 - Typical Air System
2 | Air Systems
SYSTEM
OUTLET
SYSTEM
OUTLET
2. Symbols and Subscripts
2.1 Symbols and subscripted symbols
Symbol
Description
SI
(I-P)
A
Ae
Ao
ah
C
Cd
Cn
c
D
E
ε
f
g
γ
K
L
μ
∆P
P
Ps
Psx
Pt
Ptx
Pv
p
Q
Qx
R
Re
rh
ρ
ρx
SEF
SR
sh
t
td
tw
V
v
Y
Z
~
Area
Area-Orifice Equivalent to System
Area-Nozzle with no loss
Absolute Humidity
Dynamic Loss Coefficient
Coefficient of Discharge
Coefficient of Nozzle Discharge
Speed of Sound
Duct Diameter and Equivalent Diameter
System Resistance Curve
Absolute Surface Roughness Height
Friction Coefficient
Gravity
Ratio of Specific Heats
System Effect Factor (System)
Length
Air Viscosity, Absolute
Pressure Differential
Pressure
Static Pressure
Static Pressure at Plane x
Total Pressure
Total Pressure at Plane x
Velocity Pressure
Atmospheric Pressure
Airflow Rate
Airflow Rate at Plane x
Gas Constant
Reynolds Number
Relative Humidity
Air Density
Air Density at Plane x
System Effect Factor (Fan)
System Resistance Factor
Specific Humidity (_/_ dry air)
Temperature
Dry-Bulb Temperature
Wet-Bulb Temperature
Average Velocity
Velocity - At any Point
Expansion Factor
Altitude
Is Proportional to
m2
(ft2)
m2
(ft2)
m2
(ft2)
kg/m3
(lb/ft3)
Dimensionless
Dimensionless
Dimensionless
m/s
(ft/s)
m
(ft)
Dimensionless
m
(ft)
Dimensionless
m/s2
(ft/s2)
Dimensionless
Dimensionless
m
(ft)
N-s/m2
(lbm/ft-s)
Pa
(in. wg)
Pa
(in. wg)
Pa
(in. wg)
Pa
(in. wg)
Pa
(in. wg)
Pa
(in. wg)
Pa
(in. wg)
Pa
(in. Hg)
m3/s
(cfm)
m3/s
(cfm)
J/kg-K
(ft-lb/lbm-°R)
Dimensionless
%
(%)
kg/m3
(lbm/ft3)
kg/m3
(lbm/ft3)
Pa
(in. wg)
m-4
(ft-4)
kg/kg dry air
(lb/lb) dry air
°C
(°F)
°C
(°F)
°C
(°F)
m/s
(ft/min)
m/s
(ft/min)
Dimensionless
m
(ft)
Dimensionless
2.2 Subscripts
Subscript
Definition
Subscript
Definition
a
b
c
d
E
F
Element a
Element b
Element c - Combined
Discharge
Plane of System Entry
Fan
n
O
x
x,x'
1
2
Reference to Nozzle
Plane of System Outlet
Plane 0, 1, 2,...as appropriate
Between Planes x and x'
Plane of Fan Inlet
Plane of Fan Discharge
3. Properties of Air
Atmospheric air is a mixture of several gases, water
vapor, and impurities. The relative amounts of the
important constituents for dry, sea level air are given
in Table 3.1. This table may be considered
representative of air at any altitude.
Table 3.1 - Dry Air Composition, Fraction
Component
Volume
Weight
Nitrogen
0.7809
0.7552
Oxygen
0.2095
0.2315
Argon
0.0093
0.0128
Carbon Dioxide
0.0003
0.0004
Also slight traces of neon, hydrogen, helium,
krypton, ozone and others
Although the gas composition of air can be
considered essentially constant, the amount of water
vapor contained in the air can vary greatly. The
properties of moist air are dependent upon the
relative amount of water vapor and dry air, therefore,
in defining the properties of moist air, this relative
amount must be defined (see Section 3.1.5
Humidity).
The impurities in the air are of various forms, but
basically can be divided into two categories: a)
particulates which can be either solid or liquid, and b)
mixtures, which can be either gas or vapor. The
distribution of these impurities is not uniform on an
atmospheric scale, but can be considered uniform for
the purposes of air system design. Since air is a
mixture of several gases, the behavior of air under
varying conditions can be best understood by first
reviewing the behavior of pure gases.
3.1 Properties of gases
A gas may be defined as a compressible substance
which has no free surfaces and occupies all portions
of its container. The important properties of an ideal
gas are listed below.
3.1.1 Density. The density of a gas is defined as the
total mass of the molecules in a unit volume. In the SI
system density is given in kilograms per cubic meter
(kg/m3); in the I-P system, density is given in pounds
per cubic foot (lbm/ft3).
For purposes of uniformity, standard air has been
defined as air with a density of 1.2 kg/m3 (0.075
lbm/ft3) and an absolute viscosity of 18.19 × 10-6
4 | Air Systems
N-s/m2 (1.222 × 10-5 lbm/ft-s). This is substantially
equivalent to air at a temperature of 20°C (68°F),
50% relative humidity, and a barometric pressure of
101 kPa (29.92 inches mercury) at sea level. The
ratio of specific heats, (γ), is taken to be 1.4, which is
the expected value for a perfect diatomic gas.
The temperature and barometric pressure of
atmospheric air vary widely with weather conditions
and geographical location, most noticeably altitude.
In order to simplify design, standard atmospheric
conditions have been defined which give the variation
of atmospheric pressure, temperature, and,
therefore, density with altitude. Annex B lists these
variations.
3.1.2 Pressure
In an air system, pressure is the force exerted by the
air molecules on the surfaces which make up the
system. Since air molecules are always in motion,
they continuously collide with other air molecules or a
solid surface. All these collisions are considered to be
perfectly elastic and, in the case when a molecule
strikes a surface, the surface experiences a force
equal and opposite to the time rate of change of
momentum of the rebounding molecule. This force
causes the gas to exert an overall pressure on an
immersed body and this force per unit area is
referred to as the pressure. In air system work, the
units of pressure are given in terms of force per unit
area. The unit of measure for pressure in the SI
system is the Pascal (Pa); in the I-P system the units
are inches of water gauge (in. wg).
3.1.3 Temperature
3.1.3.1 Thermal relationships
The kinetic energy of gas molecules increases with
increasing temperature. The important effects of this
fact are stated in Boyle's Law and Charles' Law,
which state that the volume of a perfect gas varies
inversely with absolute pressure and directly with
absolute temperature, respectively. The total effect is
more properly stated by the equation of state:
PV = mRT
Eq. 3.1-1
or
P = ρRT
Where:
P = Pressure
V = Volume
m = mass
R = Gas Constant
T = Absolute Temperature
ρ = m/V = density
Eq. 3.1-2
In the design of most air systems, it is acceptable to
assume that the gas is incompressible, therefore, the
air density may be considered constant, and
therefore, the absolute pressure and absolute
temperature are directly proportional.
3.1.3.2 Dry-bulb, wet-bulb and dew point
temperature. Unless otherwise specified, the
temperature of an air-water vapor mixture is that
temperature which is indicated by an ordinary or drybulb thermometer. This dry-bulb temperature is the
temperature of both the air and the water vapor in the
mixture. The wet-bulb temperature may be
determined by exposing a wetted bulb in a moving
air-water vapor mixture until equilibrium is obtained.
The wet-bulb temperature will be lower than the drybulb temperature as long as evaporation continues. If
no evaporation is possible, the mixture is saturated
and the wet and dry-bulb temperatures for this
condition will be identical. The dew point temperature
of an air-water vapor mixture is the saturation
temperature corresponding to the absolute humidity
of the mixture. The dew point temperature may also
be considered as that temperature at which
condensation begins when the mixture is gradually
cooled.
3.1.4 Viscosity. A non-perfect gas, such as air, is
capable of exerting a force parallel to the surface of a
body which is moving with respect to the gas. The
magnitude of the force parallel to the surface is used
to define an important property of non-perfect gases
- viscosity. The effects of viscosity on the behavior of
real gases cause resistance to flow; the resistance is
proportional to the velocity gradients which exist in
the gas. The absolute viscosity (µ) is defined as the
shearing stress for a unit rate of change of velocity.
The absolute viscosity has units of newton-sec per
meter squared (N-s/m2) in the SI system and pound
mass per foot-second (lbm/ft-s) using I-P units.
3.1.5 Humidity. The density of atmospheric air is
also a function of the humidity. Although the change
in density due to a change in humidity is not large, it
is often significant and air system designers should
be aware of these changes. Remember that
increasing humidity lowers the density since water
vapor is lighter than dry air. The density of saturated
air for various barometric and hygrometric conditions
is shown in Annex C.
Partially saturated air contains vapor that is
superheated, that is, the temperature of the mixture
and, therefore, that of the vapor is higher than the
saturation temperature for the existing vapor
pressure.
The relative humidity (rh) of an air-water vapor
mixture is defined as the ratio of the vapor pressure
existing compared to the vapor pressure at saturation
for the same dry-bulb temperature. This is also equal
to the ratio of the mole fractions under the same
condition. Relative humidity is always expressed as a
percent.
Specific humidity (sh) is the actual mass (weight) of
the water vapor existing per unit mass (unit weight) of
dry air or gas. Absolute humidity (ah) may be
expressed in kilograms (pounds) of water vapor per
cubic meter (cubic foot) of mixture. The humidity of
an air-water vapor mixture is often expressed by
giving either relative humidity or a wet-bulb
depression.
4. Airflow
The flow of any fluid between two points is caused by
the existence of a pressure differential between the
two points. It is the purpose of this section to explain
the parameters that may affect the flow of a gas
between two points.
4.1 Flow conditions
Most air systems are designed in the incompressible
range. Where compressibility is a factor, Mach
number and Reynolds number must be considered.
The magnitude of these parameters gives an
indication of the effects which can be expected from
the deviations in the non-perfect gas behavior from
that of a perfect gas.
4.1.1 Mach number. Mach number, for our purposes
here, is the ratio of the velocity of an airstream to the
speed of sound in that airstream.
Mach number = V/c
Where:
V = velocity of air, m/s (ft/s)
c = speed of sound in air, m/s (ft/s)
The speed of sound is a function of temperature and
is the speed at which very small pressure
disturbances are propagated throughout the gas. The
speed of sound is proportional to the square root of
the absolute temperature, and for standard air is
approximately 345 m/s (1130 ft/s). If the Mach
number is small and no large static pressure changes
are introduced by mechanical means, the flow may
be considered incompressible, that is, the density is
everywhere constant. Air can be considered
incompressible if the fan total pressure rise is less
than 2980 Pa (12 in. wg).
Air Systems | 5
4.1.2 Reynolds number. The ratio of the inertia force
to the viscous force caused by changes in velocity
within the fluid element is known as the Reynolds
number.
⎛ρ⎞
Re = DV ⎜ ⎟
⎝μ⎠
Eq. 4.1-1A SI
⎛ ρ ⎞
Re = DV ⎜
⎟
⎝ 60 μ ⎠
Eq. 4.1-1A I-P
friction drag, and, for streamlined bodies closely
aligned with the flow, represents the entire drag force.
For blunt bodies, which may be streamlined bodies at
large angles to the flow, profile drag exists. Profile
drag is caused by the inability of the flow, due to its
viscous effects, to follow the body shape. The
inability to follow the body shape creates a wake of
very turbulent flow which in effect creates the profile
drag force. These wake effects are the predominant
cause of flow losses in systems.
⎛ DV ⎞
=⎜
⎟
⎝ γ ⎠
For standard air:
Re = 65970.3DV
Eq. 4.1-1B SI
Re = 102.3DV
Eq. 4.1-1B I-P
Figure 4A - Skin Friction Drag
Where:
D = Any convenient reference dimension, m (ft)
V = Velocity, m/s (ft/min)
μ = Absolute viscosity, N-s/m2 (lbm/ft-s)
γ = Kinematic viscosity, m2/s (ft2/s)
ρ = Density, kg/m3 (lbm/ft3)
For flow about immersed bodies, D is normally taken
as the length of body in the direction of flow. In ducted
flow, D is normally taken as the diameter of the duct;
in unducted flow, D is normally taken as the diameter
of the opening through which the flow passes.
For a fan, D is equal to the impeller tip diameter and
is only proportional to conventional Reynolds
numbers. The Reynolds number provides a
convenient non-dimensional means of comparing
two flows.
4.2 Flow about immersed bodies
If a solid body is immersed in a flowing stream of a
gas, the direction of flow of the gas will be parallel to
the surface of the solid body. The changes in the
direction of the molecules close to the body exert
forces on the body which when taken over the entire
body, are perpendicular to the direction of the gas
flow.
A non-perfect gas will also exert a force parallel to the
direction of the velocity, due to the viscosity of the
gas. This force, usually called drag, is due to two
effects. The first is the shearing force set up within
the molecules of the gas resulting from the molecules
decelerating from the gas velocity to zero velocity
when in contact with the body. This is called skin
6 | Air Systems
Figure 4B - Profile Drag
Figures 4A and 4B illustrate skin friction drag and
profile drag.
4.3 Ducted flow
When air flows through a duct of constant crosssection, the average velocity remains constant and is
parallel to the center line of the duct. Due to friction,
the velocity at the duct wall is zero and the average
velocity profile can be defined as either of two
conditions:
a) Laminar Flow: Flow in which the air velocity
vectors are parallel to the duct wall. This type of
flow is described as smooth.
b) Turbulent Flow: Flow in which air velocity vectors
at various points across the duct are at various
angles, up to and including reverse flow.
Except for extremely low air velocities, laminar flow
does not exist and all duct flow involving air can be
considered to be in the transition region between
laminar and fully turbulent flow. The transfer of
energy from the high velocity section in the center of
the duct to the low velocity section near the duct wall
causes a marked resistance to the flow. This
resistance varies linearly with the length of the duct
and approximately with the square of the average
velocity in the duct. The resistance is also a function
of the Reynolds number of the flow, which is
calculated using the average velocity in the duct, the
duct diameter, and the surface roughness of the duct
wall.
The velocity profiles in a duct system for fully
developed flow will vary depending on whether the
flow is laminar or turbulent and the degree of duct
roughness. Velocity profiles of various flow conditions
are shown in Figure 4C.
The absolute velocity of the air stream will vary
substantially over the cross-sectional duct area, but
for duct systems the velocity used for determining the
velocity pressure is always the average velocity given
by:
V average = Q/A
Eq. 4.3-1
Where:
V = Velocity, m/s (ft/min)
Q = Flow rate, m3/s (cfm)
A = Area of the cross-section where the flow occurs,
m2 (ft2)
The duct velocity profiles shown in Figure 4C are
uniform along the length of the duct and symmetrical
around the center line. Where there are disturbances
in the ducts, such as turns, expansion or contraction,
etc., the velocity profile across the duct can become
very asymmetrical as shown in Figure 4D.
The flow will return to a normal velocity profile after a
disturbance if there is sufficient length of straight duct
to allow the velocity distribution to regain uniformity. A
minimum of 2½ equivalent duct diameters of straight
duct is required to attain a normal velocity profile for
velocities of 12.7 m/s (2500 ft/min) or less. Add one
duct diameter for each additional 5.08 m/s (1000
ft/min). See AMCA Publication 201-90, Fans and
Systems.
4.4 System losses
The losses in total pressure for flow through a system
are caused by two factors: friction losses due to
viscosity as the air flows along the surface of ducts
and other system elements, and dynamic losses
due to the turbulent wake caused by changes in
direction and separation of the flow around obstructions.
In addition to the losses in total pressure in a system
caused by friction losses and dynamic losses, there
are losses due to System Effects.
System Effects occur because of the differences
between the fan inlet and outlet connections to the
installed system and the standardized connections
used in laboratory tests to obtain fan performance
ratings. AMCA Publication 201, Fans and Systems,
gives specific details on System Effects related to
fans. System Effects related to series system
elements are covered further in Section 4.5 of this
publication.
4.4.1 Duct friction losses. In the normal range of air
systems for HVAC and industrial applications, the
flow falls into the transition region between laminar
flow and complete turbulent flow. In this region the
losses due to friction are a function of Reynolds
number and the relative roughness of the duct wall.
The pressure loss in the transition region will vary at
slightly less than the square of the velocity. The
pressure loss due to friction for flow in ducts may be
calculated from the Darcy-Weisbach equation:
∆Pt = f(L/D) Pv
Eq. 4.4-1
Where:
∆Pt =
f
=
L =
D =
Pv =
Total pressure loss due to friction, Pa (in. wg)
Friction factor, dimensionless
Length of duct, m (ft)
Diameter of pipe, m (ft)
Velocity pressure, Pa (in. wg)
In the transition flow range, the value of the friction
factor cannot be calculated directly. It can be
obtained from the Moody diagram or by iterative
solution of the Colebrook equation. See the ASHRAE
Handbook: Fundamentals, chapter on Duct Design,
for a more complete discussion of duct friction losses.
The Moody diagram, Figure 4E, shows the
relationship of the friction factor, Reynolds number
and duct roughness (ε) in meters (feet). Most
applications are in the transition region between
laminar and full turbulent flow conditions.
Using duct friction charts (see Annex D) is the most
common method of determining friction losses.
These charts are based on ducts having average
roughness and standard air density. Correction
factors must be applied for ducts having different
roughness, and for variations in air density and
viscosity.
Air Systems | 7
r
LAMINAR
TURBULENT
SMOOTH Re = 107
SMOOTH Re = 105
0
0.5
ε = 0.03D ROUGH
ε = 0.008D ROUGH
1.0
1.5
v
V
D = Duct Diameter
ε = Duct Roughness
Re = Reynolds Number
v = Velocity at any Point
V = Average Velocity
r = Radius
Figure 4C - Velocity Profiles in a Round Duct for Various
Reynolds Numbers and Duct Roughness
Figure 4D - Changing Velocity Profiles
8 | Air Systems
2.0
0.10
0.09
0.08
0.05
0.07
0.04
0.06
0.03
0.05
0.02
0.010
0.008
0.006
RO
0.03
UG
0.004
H
N
PE
DE )
Re 9a
H (2
IT Eq.
W
FRICTION FACTOR, f
0.04
0.002
NC
DE
RELATIVE ROUGHNESS, ε/D
0.015
FULLY ROUGH (EQ 18)
Eq. (29a)
E
0.02
0.0010
0.0008
0.0006
Eq. (27)
0.0004
SMOOTH PIPE
Eqs. (28a) and (28b)
0.015
0.0002
LAMINAR
TRANSITION
REGION
TURBULENT
0.00010
0.00005
0.010
0.009
0.008
103
2
3
5
104
2
3
5
105
2
3
5
106
2
3
5
107
2
3
5
0.00001
108
REYNOLDS NUMBER, Re
Figure 4E - Moody Diagram
Reprinted by permission of the American Society of Heating, Refrigerating and Air-Conditioning Engineers, Atlanta, Georgia, from the 1993
ASHRAE Handbook-Fundamentals. (Moody 1944). Values on the chart are the same for both the SI and I-P systems. Equation numbers refer
to equations in the source document.
Air Systems | 9
Loeffler1 has developed simplified equations for the
friction factor in the normal range of flow conditions
found in industrial and HVAC air systems. The
equations provide for direct calculation of duct friction
total pressure losses. These equations yield results
that are accurate within ±5% and are conservative
over most of the range of flow.
For aluminum ducts, medium smooth:
ε = 0.0000457 m (0.00015 ft)
Correction factors for density and viscosity variations
need to be applied for conditions other than standard
air. See Annexes B and E.
4.4.2 Dynamic losses. Dynamic losses occur where
there are changes in velocity or direction in the air
system and are due primarily to the profile drag.
Dynamic losses are proportional to the square of
velocity, and therefore, are proportional to the
velocity pressure.
∆Pt ~ V 2 ~ Pv
⎛ LQ1.863 ⎞
Pt loss = a ⎜ 4.93 ⎟
⎝ D
⎠
Where:
a = 1.764 × 10-2
Eq. 4.4-2 SI
a = 4.816 × 10-9
Eq. 4.4-2 I-P
Dynamic pressure loss data are given in a number of
forms such as pressure loss for given volume or
velocity, equivalent length of duct, or velocity
pressure multiplier, and is available from
manufacturers' data and handbooks such as:
ASHRAE, the Industrial Ventilation Guide and
SMACNA.
Except at duct exits, dynamic losses occur along
some length and cannot be separated from friction
losses. For practical purposes the dynamic losses
are assumed to be concentrated at one point and the
friction losses are included as part of the duct friction.
Dynamic loss coefficients for duct fittings are based
on zero length. For friction loss calculations, the
centerline length of the duct fitting is taken as the
length of the fitting.
For galvanized steel ducts, average:
ε = 0.0001524 m (0.0005 ft)
⎛ LQ1.921 ⎞
Pt loss = a ⎜ 5.066 ⎟
⎝ D
⎠
Where:
a = 1.717 × 10-2
Eq. 4.4-3 SI
a = 3.534 × 10-9
Eq. 4.4-3 I-P
For fiberglass ducts or lined ducts, fabric and wire
flexible ducts (wire covered with fabric), medium
rough,
ε = 0.00091443 m (0.003 ft)
⎛ LQ1.965 ⎞
Pt loss = a ⎜ 5.208 ⎟
⎝ D
⎠
Where:
a = 2.093 × 10-2
Eq. 4.4-4 SI
a = 3.64 × 10-9
Eq. 4.4-4 I-P
4.4.3 Dynamic loss coefficient. There are two
common methods of expressing dynamic losses.
These are: 1) The equivalent length of duct method,
and 2) the loss coefficient method. The equivalent
length of duct method replaces the dynamic loss of
fittings (elbows, tees, branches, etc.) with a length of
duct that will have an equivalent loss. The equivalent
length of duct for all of the dynamic losses are added
to the straight duct length. A friction chart showing the
loss is then used to determine the total loss in the
system.
The dynamic loss coefficient method is based on the
fact that all losses in a system are functions of the
velocity pressure and can be calculated by a
corresponding dynamic loss factor multiplied by the
velocity pressure. These losses are added to the
straight duct friction loss to determine the total loss in
the system.
Where:
Pt loss = Total pressure loss, Pa (in. wg)
Q
= Flow rate, m3/s (cfm)
D
= Duct diameter, m (ft) (or equivalent diameter
of rectangular ducts) D equivalent = (4ab/ )0.5
where a and b are the sides in m (ft)
L
= Duct length in m (ft)
10 | Air Systems
The dynamic loss coefficient method is preferred
because it is usually quicker and offers the
advantage of faster recalculation when other branch
duct sizes are tried.
Dynamic losses are proportional to the velocity
pressure occurring in the system element and,
therefore, the pressure loss in the fitting can be
related to the velocity pressure by use of a dynamic
loss coefficient.
The dynamic loss coefficient Co is defined as:
Co = ∆Pt/Pvo
Eq. 4.4-5
Where:
Co = Dynamic loss coefficient, reference to section
o, dimensionless
∆Pt = Dynamic pressure loss, Pa (in. wg)
Pvo = Velocity pressure at section o, Pa (in. wg)
The coefficient relates the pressure loss in the
element to the velocity pressure at a given crosssectioned area of the element.
The pressure loss of duct system elements with
known dynamic loss coeffients can be calculated by:
∆Pt = Co Pvo
Eq. 4.4-6
Where there are changes in area or divided flow in
the fitting, the designer must be careful to use the
proper area as noted in the loss tables for the
determination of the velocity pressure to be used with
the dynamic loss coefficient. The ASHRAE
Handbook: Fundamentals, Chapter on Duct Design,
provides a detailed discussion of the dynamic loss
coefficient and tables for coefficients of many
common duct elements.
4.5 System Effects
Additional losses can occur in air systems because of
the physical relationship of various elements in the
system. These System Effect losses occur because
of the difference between the way the performance of
the element was determined by testing and the way
the element is actually installed in the system.
4.5.1 Fan System Effects. Fan System Effects occur
because of the difference in inlet and outlet
conditions under laboratory test conditions and the
inlet and outlet conditions as the fan is installed in the
system. Detailed information on Fan System Effects
is contained in AMCA Publication 201, Fans and
Systems. The System Effect is accounted for as a
pressure loss which must be included with the other
system losses. The sum of the pressure losses is
then used as the basis for selecting the fan.
4.5.2 Element System Effects in series. System
Effects for other air system elements occur when two
or more elements are in close proximity to one
another. Loss coefficients for duct fittings, coils,
filters, dampers, etc., are determined with a sufficient
length of straight duct (normally 10 diameters) ahead
of the element to allow for a normally distributed
velocity profile entering the element, and a sufficient
length of straight duct (normally 10 diameters)
downstream from the element to allow a normally
distributed velocity profile to be re-established. When
two elements, such as elbows, or an elbow and a
damper, are placed close together the air entering
the second element will be highly turbulent and
asymmetrical in profile, causing a higher loss than
expected. In addition, any static regain occurring
downstream of the first element would also be lost
when sufficient length of straight duct is not present.
To illustrate System Effects for duct elements, the
loss coefficients for a single 90° elbow and two
elbows in series are shown in Figure 4F.
In the case of the two elbows in series, the difference
between twice the loss of the single elbow and the
actual combined loss is the System Effect. The
System Effect varies substantially depending on how
close the two elbows are to each other.
Similar effects can be expected when any system
elements are in close proximity. The amount of the
System Effect will vary over a rather wide range
depending upon the physical characteristics of each
element and their relationship to each other in the
system.
Very little actual data is available on System Effects
of various combinations of system elements, and the
system designer must, of necessity, estimate the
System Effects. The following tables for estimating
System Effects, and Equation 4.5-1, are given as a
guide to the designer. Actual data should be used
whenever it is available.
Cc = (Ca + Cb) K
Eq. 4.5-1
Where:
Cc = Loss coefficient of combined elements,
dimensionless
Ca = Loss coefficient of element a, dimensionless
Cb = Loss coefficient of element b, dimensionless
K = System Effect Factor, dimensionless
1. Loeffler, J. J., Simplified Equations for HVAC Duct Friction Factors, ASHRAE Journal, January, 1980
Air Systems | 11
ONE ELBOW
C1 = 1.15*
FLOW
Figure 4G should be used for elements in series
where the flow is straight through, while Figure 4H
should be used when a turning element (elbow, etc.)
is involved. These tables have been developed on
the basis of limited data and are intended only as a
guide. Actual System Effects may vary from the
values shown. See sample calculation in Section 5.8
TWO ELBOWS IN SERIES
a
C2 = VARIES
FLOW
b
RESISTANCE
ELEMENTS
D
L
L
D
L/D
1.0
2.0
4.0
10.0
C2 Loss
coefficient
for two
elbows in
series**
2.63
4.18
3.08
2.45
2 C1 Two
times the
loss
coefficient
of a single
elbow
System
Effect
(Difference)
L/D
.50
1
2
3
4
5
7.5
10
K
1.5
1.4
1.3
1.2
1.15
1.1
1.05
1.0
From SMACNA Duct Design Manual
Figure 4G - Element System Effects for
Straight Through Flow
2.30
+0.33
14.3%
+81.7%
1.88
+33.98%
0.78
a
b
+6.5%
0.15
L
D
* From I. E. Idelchik, Handbook of Hydraulic Resistance, 3rd
Edition, 1993, p.365, Hemisphere Publishing Company.
** ibid, p.375
Figure 4F - System Effect of Duct Elements
L/D
.50
1
2
3
4
5
7.5
10
K
1.0
1.08
1.75
1.5
1.30
1.2
1.1
1.0
Figure 4H - Element System Effect for
Turning Elements
12 | Air Systems
5. The Flow System
5.1 Concepts of pressure
The flow of air between two zones (or spaces) is due
to a pressure difference between the two zones. This
pressure difference forces the air to flow from the
high pressure zone to the low pressure zone.
Ductwork is used in most air systems to convey the
air from one zone to another. The quantity of air Q in
m3/s (cfm) that will flow is equal to the cross-sectional
area A of the duct in m2(ft2) times the air velocity V in
m/s (ft/min).
Q = AV
Eq. 5.1-1
ρ is equal to the air density in kg/m3 (lbm/ft3), and
standard air density equals 1.2 kg/m3 (0.075 lbm/ft3).
Pv is always positive and this pressure is always
exerted in the direction of airflow. Air confined within
a duct or a tank, whether in motion or not, creates
another kind of pressure which exerts itself in all
directions at once, including perpendicular to the
walls of the enclosure. This is known as static
pressure Ps. Static pressure is negative when it is
below atmospheric pressure, and positive when
above atmospheric pressure. Total pressure Pt at any
point in an air system is equal to the algebraic sum of
the static pressure Ps and velocity pressure Pv. Pt can
be either positive or negative, depending on its
components.
The air traveling at a given velocity V in m/s (ft/min)
will create a velocity pressure Pv in Pa (in. wg). The
velocity pressure in these terms is:
Pt = Ps + Pv
Pv = 0.5ρV2
Eq. 5.1-2A SI
Pv = ρ(V/1096)2
Eq. 5.1-2A I-P
Of prime concern in air system design is the
relationship of these pressures internal to the air
system. The significance of these pressures can be
demonstrated on the next several pages in Figures
5A, 5B, and 5C.
And for standard air:
Pv = 0.6V2
Eq. 5.1-2B SI
Pv = (V/4005)2
Eq. 5.1-2B I-P
Eq. 5.1-3
5.2 Examples of pressures in duct systems
Air Systems | 13
In Figure 5A, the sealed length of duct has a static pressure of 345 Pa (1.39 in. wg) above atmospheric pressure.
Since there is no airflow, velocity pressure is equal to zero. The total pressure Pt can then be calculated according
to equation 5.1-3.
Pt
SI
= Ps + Pv
= 345 + 0
= 345 Pa
Pt
TOTAL PRESSURE 345 Pa (1.39 in. wg)
I-P
= Ps + Pv
= 1.39 + 0
= 1.39 in. wg
STATIC PRESSURE
345 Pa (1.39 in. wg)
Figure 5A - Sealed System
14 | Air Systems
=
VELOCITY PRESSURE
0 Pa (0 in. wg)
In Figure 5B, with the duct open and a fan placed at one end blowing air through the duct, we find both static
pressure and velocity pressure as illustrated by the water gauge. The total pressure is the sum of velocity pressure
and static pressure.
TOTAL PRESSURE 345 Pa (1.39 in. wg)
STATIC PRESSURE
97 Pa (0.39 in. wg)
=
AIRFLOW
VELOCITY PRESSURE
248 Pa (1.00 in. wg)
20.3 m/s
(4005 FPM)
In this illustration, the static pressure will be above atmospheric pressure and the total pressure is numerically
greater than either static or velocity pressure.
Figure 5B - Positive Pressure System
Air Systems | 15
In Figure 5C, a fan is placed at the end of the duct and draws air through the duct. In this case, the static pressure
is below atmospheric pressure. In both Figures 5B and 5C, the total pressure rise across the fan is 345 Pa (1.39
in. wg); therefore, the energy used in both systems is equal.
TOTAL PRESSURE
STATIC PRESSURE
-97 Pa (-0.39 in. wg)
-345 Pa (-1.39 in. wg)
=
VELOCITY PRESSURE
248 Pa (1.00 in. wg)
20.3 m/s
(4005 FPM)
AIRFLOW
Figure 5C - Negative Pressure System
16 | Air Systems
5.3 Conservation of energy
Therefore, to simplify Bernoulli's Theorem:
The well known Bernoulli Theorem, frequently used
in the flow of fluids, states the law of conservation of
energy for fluid systems. In its simpler form for
airflow:
Ps1 + Pv1 = Ps2 + Pv2 + losses1,2
(V12/2g) + (P1/ρg) + Z1 = (V22/2g) + (P2/ρg)
+ Z2 + losses1,2
Where:
V 2/2g =
Kinetic energy or velocity pressure (Pv)
x
Px/ρg = Potential energy or static pressure (Ps)
Zx
= The elevation (normally ignored in fan
systems with minimal changes in
elevation) may need to be evaluated when
the system elevation change causes a
noticeable change in air density.
Pt
PS
PT. #1
PV
PRESSURE
Pa (INCHES W.G.)
Simply stated, the sum of static pressure and velocity
pressure at any point in a flow system is equal to the
sum of static pressure and velocity pressure at any
other point in the system, plus any losses in pressure
occurring between the two points.
Although it is very important to recognize that the
sum of static and velocity pressure remains constant
at any point in the system when disregarding losses,
it is the losses that are of real importance in the
design and function of air moving systems.
To illustrate Bernoulli's Theorem, Figure 5D shows a
typical venturi system commonly used to measure
fluid flow.
Pt
PS
PV
PT. #2
Pt
PS
PV
Eq. 5.3-1
Pt
PT. #3
PS
PV
AIRFLOW
Pt
PS
PV
Figure 5D - Venturi System
Air Systems | 17
For the purposes of illustration, assume this is a
perfect device with no losses. The velocity pressure,
static pressure, and total pressure at each point in the
system are shown on the gauges as well as on the
lower portion of the illustration. Part of the static at
point 1 is converted to velocity pressure as the fluid
is accelerated in the contracted flow area in the neck
of the venturi. At point 2, in the neck of the venturi,
the velocity pressure reaches its maximum and the
static pressure is reduced, but the total pressure
remains constant. As the flow area is gradually
increased to point 3, the velocity is slowed down to
the original velocity and the velocity pressure is
reduced to its original value. The static pressure
increases back to its original value also, while the
total pressure remains constant.
In this illustration, part of the static pressure at point
1 is converted to increased velocity pressure at point
2 and the velocity pressure is then converted back
into static pressure at point 3.
The conversion of the velocity pressure into static
pressure by reducing the velocity is known as static
regain and it is very important to understand this
phenomenon in the design of flow systems.
In air systems, changes in velocity often occur as the
air flows through the various elements comprising the
system. Some total pressure loss will occur any time
the velocity in the system is increased or decreased.
The magnitude of these losses is dependent on the
physical characteristics of the system element in
which the velocity change takes place. Decreases in
velocity occurring at abrupt enlargements in area
result in total pressure losses approaching one
velocity pressure.
Figure 5E illustrates air flowing from a large plenum
through a long radius flow nozzle and discharging to
atmosphere.
If the plenum is very large relative to the nozzle
diameter, the velocity in the plenum will approach
zero, so the total pressure at point 1 (Pt1) will be equal
to the static pressure (Ps1). Assuming a perfect fluid
and no losses occurring in the nozzle, the static
pressure in the plenum (point 1) is totally converted
to velocity pressure at the discharge of the nozzle
(point 2).
At point 2, just beyond the discharge end of the
nozzle, the static pressure (Ps2) will be zero (at
atmospheric pressure) and the total pressure (Pt2) will
be equal to the velocity pressure of the air stream.
Applying Bernoulli's Equation for points 1 and 2
gives:
Ps1 + Pv1 = Ps2 + Pv2
Since Pv1 is equal to 0 and Ps2 is equal to 0, the
equation reduces to:
Ps1 = Pv2
In this case, the energy of the static pressure in the
plenum is totally converted to the velocity energy at
the discharge of the nozzle. It should also be noted
that the velocity pressure is totally lost as the
airstream discharges to atmosphere.
Using Equation 5.1-2A (as reduced above) the
relationship between static pressure and the throat
velocity can be established in the above example.
Using this in Equation 5.1-2A gives:
V2 = (2Ps1/ρ2)0.5
Eq. 5.3-2A SI
V2 = 1096(Ps1/ρ2)0.5
Eq. 5.3-2A I-P
Or where: Ps1 = Pv2
Then for standard air, the equations would be:
V2 = 1.29(Ps1)0.5
Eq. 5.3-2B SI
V2 = 4005(Ps1)0.5
Eq. 5.3-2B I-P
The throat velocity pressure would be:
POINT 1
POINT 2
Figure 5E - Airflow Through a Nozzle
Pv2 = 0.5ρ2V22
Eq. 5.3-3A SI
Pv2 = (V2/1096)2ρ2
Eq. 5.3-3A I-P
Where:
Ps1 = Pv2
18 | Air Systems
Then, for standard air, the equation would be:
Pv2 = 0.6V 2
Eq. 5.3-3B SI
Pv2 = (V2/4005)2
Eq. 5.3-3B I-P
In the above example, the quantity of flow would be
dependent on the area of the nozzle discharge and
the velocity in the throat as given by the equation:
Q2 = V2A2
Where:
Q = Airflow rate, m3/s (cfm)
V = Velocity, m/s (ft/min)
A = Area, m2 (ft2)
In practical system design both Cn and Y approach
unity and can normally be neglected. However, these
factors are important when measuring fan
performance in accordance with ANSI/AMCA
Standard 210.
In Figure 5F, the nozzle is replaced by a sharp edged
orifice. The flow through the orifice tends to neck
down to a flow area smaller than the orifice diameter.
The point at which the flow area reaches its minimum
is called the vena contracta. The flow through the
orifice is given by the equation:
Q = Cd(2Ps/ρ)0.5A
Eq. 5.3-6 SI
Q = Cd1096(Ps/ρ)0.5A
Eq. 5.3-6 I-P
Where:
Substituting for V2 from Equation 5.3-2A and 5.3-2B
gives:
Q = (2Ps1/ρ2)0.5A2
A = Area of orifice, m2 (ft2)
Cd = Coefficient of discharge, dimensionless
Eq. 5.3-4A SI
VENA CONTRACTA
Q = 1096(Ps1
/ρ )0.5A
2
2
Eq. 5.3-4A I-P
And for standard air:
Q = 1.29(Ps1)0.5A2
Eq. 5.3-4B SI
Q = 4005(Ps1)0.5A2
Eq. 5.3-4B I-P
The flow nozzle in the above illustration is the basis
for airflow measurement in the ANSI/AMCA Standard
210 duct nozzle and chamber nozzle methods of
airflow measurement. The ANSI/AMCA Standard 210
nozzle very closely approaches perfect flow
conditions with almost zero losses. There are,
however, some losses which vary with Reynolds
number as well as the effect of the compressibility of
the gas which must be accounted for. The Equation
5.3-4A must be modified by adding factors for the
losses and compressibility.
The equation of flow becomes:
Q = CnY(2Ps/ρ)0.5A
Q = CnY1096(Ps/ρ)0.5A
Where:
Q = Airflow, m3/s (cfm)
Cn = Nozzle coefficient, dimensionless
Y = Expansion factor, dimensionless
A = Area of nozzle, m2 (ft2)
Eq. 5.3-5 SI
5.3-5 I-P
Figure 5F - Airflow Through an Orifice
In this case, the coefficient of discharge (Cd)
accounts for the reduction in flow area of the vena
contracta and losses occurring in the flow system.
The value of the coefficient of discharge is dependent
on the Reynolds number of the flow system. Where
sharp edged orifices are used in the duct system, the
coefficient of discharge is also dependent on the ratio
of orifice diameter to duct diameter.
As in the previous example, all of the velocity
pressure is lost as the airstream discharges to
atmosphere. Where ducts or plenums are used on
the discharge of nozzles or orifices, a portion of the
velocity pressure at the nozzle discharge is regained
as static pressure as the velocity returns to a normal
distribution.
Air Systems | 19
If we plot the flow through these elements versus
static pressure as shown in Figure 5G, we obtain a
squared curve which is typical of flow versus
pressure for any constant system.
Resistance Curve (see Figure 5H) the point of
operation is at the intersection of the fan performance
curve and the System Resistance curve.
PRESSURE
PRESSURE
SYSTEM
RESISTANCE
P = KQ2
POINT OF
OPERATION
FAN PRESSURE
CURVE
FLOW
FLOW
Figure 5G - Typical System Resistance Curve
These curves are called System Resistance curves,
and define the relationship of flow versus pressure
for any system with constant resistance. System
Resistance curves are defined by the following
equations:
Q = (2Ps/ρ)0.5Ae
Eq. 5.3-7A SI
Q = 1096(Ps/ρ)0.5Ae
Eq. 5.3-7A I-P
Then for standard air:
Q = 1.29(P )0.5A
e
Eq. 5.3-7B SI
Q = 4005(Ps)0.5Ae
Eq. 5.3-7B I-P
s
Where:
Q = Airflow, m3/s (cfm)
Ps = Pressure, Pa (in. wg)
ρ = Gas density, kg/m3 (lbm/ft3)
Ae = Area of an orifice having resistance equivalent
to the system resistance, m2 (ft2) (equivalent
orifice)
The System Resistance concept is very useful in
understanding flow in complete systems or elements
of flow systems. If we add a fan curve to the System
Figure 5H - Typical Point of Operation
5.4 Fan total and static pressure
The flow of a gas through a system of ducts and
various system elements requires energy:
a) To accelerate the air from ambient conditions at
the entry to the system
b) To overcome the losses in the system from
friction and system element shock losses
c) For the loss of energy dissipated as velocity at
the system outlet
d) To overcome any static pressure at the entry or
outlet of the system
The fan provides this energy by the increase in total
pressure from the fan's inlet to the fan's outlet. The
inlet plane of a fan is referred to as Plane 1 and the
outlet plane as Plane 2. The total pressure provided
by the fan is made up of static pressure and velocity
pressure components.
The total pressure of a fan is defined as
PtF = Pt2 - Pt1
Or:
PtF = Ps2 + Pv2 - Ps1 - Pv1
20 | Air Systems
Eq. 5.4-1
Eq. 5.4-2
Eq. 5.4-4
Both static pressure and total pressure curves of the
fan and system resistance are shown. Either set of
curves can be used for the flow rate because the
point of operation is the same. The difference in
pressure at the point of operation between the total
pressure curve and the static pressure curve is the
velocity pressure at the fan discharge (plane 2).
Fan static pressure, as defined, is a term that is
peculiar to fans and is not consistent with the normal
meaning of static pressure rise. Fan static pressure is
derived from the method of testing fans where the fan
static pressure is equal to gauge static pressure at
the fan discharge when the fan draws air from
surrounding atmosphere through a well shaped inlet.
Most fans are rated in terms of static pressure and
flow, however, fans having high discharge velocities
such as vaneaxial fans are quite often rated in terms
of total pressure. Be aware of these different
methods of rating and be certain whether fan static
pressure or fan total pressure was used to determine
the fan selection.
Special care must be used when using fan static
pressure for purposes of matching the required fan
performance against system total pressure losses.
The relationship of these pressures is covered in
detail for various systems further on in this section.
5.4.2 Fan system pressure relationships. Figure
5K shows the relationship of total pressure, static
pressure and velocity pressure for a fan with free inlet
conditions and discharging through a duct against
some system. In this example the fan is shown as
being equipped with a short inlet duct and an inlet
bell. For simplicity's sake it is assumed that there are
no losses at the inlet to the fan.
The velocity pressure of a fan is defined as:
PvF = Pv2
Eq. 5.4-3
The static pressure of a fan is defined as:
PsF = PtF - PvF
The static pressure of a fan can also be stated in
several other forms. Substituting Equation 5.4-2 for
the total pressure of the fan in Equation 5.4-4 gives:
PsF = Ps2 + Pv2 - Ps1 - Pv1 - Pv2
Eq. 5.4-5
Simplifying:
PsF = Ps2 - Ps1 - Pv1
Eq. 5.4-6
Since:
Ps1 + Pv1 = Pt1
The equation can be restated as:
PsF = Ps2 - Pt1
Eq. 5.4-7
5.4.1 Fan performance specification. The system
designer should be aware of the effect of the velocity
pressure at the outlet of the system and the velocity
pressure of the fan discharge (plane 2) on the
determination of fan total or static pressure for the
system. The net result of the fan total or static
pressure requirements at a given flow rate for the
system is the fan performance specification, which is
normally stated as flow at a specific static, or total
pressure. This statement of required fan performance
is, in reality, a statement of one point on a system
resistance or equivalent orifice curve, which then
defines the flow and pressure relationship of the
system being designed.
The actual point of operation of the combined fan and
system will be the intersection of the fan performance
curve and the system curve as shown in Figure 5J.
The total pressure prior to the entry of the fan in
Figure 5K is zero and since it was assumed that there
are no entry losses, the total pressure remains zero
until the flow is acted upon by the fan. As air enters
the fan, its velocity and the pressure due to that
velocity (velocity pressure) increases while static
pressure decreases in direct proportion. (In actual
conditions there will be some entry losses which will
be accounted for in the fan performance rating.)
Referring to Figure 5K, the fan total pressure is equal
to the total pressure at plane 2 minus the total
pressure at plane 1.
PtF = Pt2 - Pt1
Eq. 5.4-8
The fan static pressure is equal to the total pressure
at plane 2 minus the velocity pressure at plane 2.
PsF = Pt2 - Pv2
Eq. 5.4-9
The static pressure of the fan can also be stated as
the static pressure at plane 2 minus the static
pressure at plane 1 minus the velocity pressure at
plane 1.
PsF = Ps2 - Ps1 - Pv1
Eq. 5.4-10
Or, as the static pressure at plane 2 minus the total
pressure at plane 1.
PsF = Ps2 - Pt1
Eq. 5.4-11
Air Systems | 21
The actual static pressure rise across the fan from
plane 1 to plane 2 will be greater than the fan static
pressure by the amount of the velocity pressure at
the fan inlet, plane 1. The difference in the actual
static pressure rise across the fan and the fan static
pressure represents the energy required to
accelerate the air to its entry velocity. This kinetic
energy is retained by the moving air stream until
there is a change in velocity in the system or it is
dissipated at the point of discharge, and as such
does not represent a loss in total pressure until it is
discharged.
5.5 The total system
The equipment arrangement shown in Figure 5K is
typical of the test conditions for fans in ANSI/AMCA
Standard 210, and is the basis for fan performance
ratings. When fans are tested in accordance with
ANSI/AMCA Standard 210, the inlet and discharge
conditions are rigidly specified for each test method.
PtE = PtO + Pt loss E,O - PtF
When a fan is installed in an air system where the
actual entry and exit conditions are different than the
test conditions, the performance of the fan may be
altered and System Effect factors must be used to
account for the altered performance. AMCA
Publication 201, Fans and Systems describes
various System Effects and provides quantitative
data for calculating System Effect losses.
A fan provides the total pressure to move the air
through a system and the flow rate will reach a point
of equilibrium (point of operation) when the total
pressure provided by the fan equals the total
pressure losses in the system at that flow rate. A
typical system with inlet and discharge resistance is
illustrated in Figure 5L.
Applying Bernoulli's Equation at the point of entry and
the outlet:
Eq. 5.5-1
Rearranging:
PtF = Pt loss, E,O + PtO - PtE
Eq. 5.5-2
Restating in terms of Ps and Pv:
PtF = Pt loss E,O + PsO + PvO - PsE - PvE
PtF = Pt loss E,O + PvO + (PsO - PsE - PvE)
Eq. 5.5-3
Substituting from Equation 5.4-4 for PtF :
PRESSURE
AIRFOIL FAN - SWSI
PvF
SYSTEM RESISTANCE
(TOTAL PRESSURE)
PtF
PsF
SYSTEM RESISTANCE
(STATIC PRESSURE)
VOLUME FLOW RATE
Figure 5J - Constant Speed Performance Curve with System Resistance
22 | Air Systems
PsF = [Pt loss E,O + PvO + (PsO - PsE - PvE)] - PvF
is included in the total pressure loss of the discharge
element.
Eq. 5.5-4
System total pressure loss = total pressure loss
internal to the system, plus the velocity pressure loss
at the outlet(s) of the system.
Equations 5.5-3 and 5.5-4 are the general statements
of fan total pressure or fan static pressure required
for flow through a system. The Pt loss term is the loss
internal to the system from friction and shock losses.
The PvO term represents the energy loss to the
system contained in the velocity at the outlet of the
system.
The term (PsO - PsE - PvE) represents the change in fan
total pressure or fan static pressure requirements
because of the static pressure conditions existing at
the system entry or outlet and any velocity pressure
present at the system entry. Velocity and velocity
pressure generated by external sources, such as
wind, at the system entry are seldom encountered,
so the PvE term can generally be disregarded.
In Equation 5.5-4, the items enclosed in large
brackets represent the fan total pressure.
In the normal method of calculating system total
pressure losses, the velocity pressure at the outlet(s)
PLANE 2
PLANE 1
E
ENTRY
PtF = Pt2 - Pt1
PsF = PtF - Pv2
RESISTANCE
ELEMENT
OUTLET
FLOW
0
FAN
Pt
Pv1
Pt2 = PtF
Pv
+
Pv2
PtE= 0
Ps2 = PsF
0
Ps1
_
Ps
ABSOLUTE
PRESSURE
Volume Flow Rate m3/sec (cfm)
Figure 5K - Fan with Discharge Resistance
(AMCA Installation Type B)
Air Systems | 23
RESISTANCE
ELEMENT
PLANE 1
PLANE 2
RESISTANCE
ELEMENT
E
ENTRY
FLOW
O
OUTLET
FAN
Pv2
PtF
ATMOSPHERIC
PRESSURE
Pv
Pt2
PvO = PtO
Ps2
+
PtE = 0
_
0
PsF
Pt
Ps
ABSOLUTE
PRESSURE
Ps1
Pvt
PtE = PtO + Pt LOSS - PtF
PtF = Pt LOSS + PvO + (PsO - PsE - PvE)
PsF = [Pt LOSS + PvO + (PsO - PsE - PvE)] - Pv2
Figure 5L - Fan with Inlet and Discharge Resistance
(AMCA Installation Type D)
PLANE 1
PLANE 2
E
O
OUTLET
ENTRY
FLOW
FAN
Ptf = PvF
+
0
PtF = PvF + PsF
since
PsF = 0
PtF = PvF = Pv2
_
Figure 5M - Fan with No Resistance at Either the Inlet or the Discharge
(AMCA Installation Type A)
24 | Air Systems
RESISTANCE
ELEMENT
E
PLANE 1
ENTRY
FLOW
PLANE 2
FAN
OUTLET
0
ATMOSPHERIC
PRESSURE
+
PtE = 0
_
Pt FAN
Pv
PtO = Pv2
Pt
Ps
Pt1
Ps
FAN
Ps1
Pv
ABSOLUTE
PRESSURE
PtF = Pt LOSS + PvO
Where:
PsF = Pt LOSS + PvO - Pv2
since PvO = Pv2
PsF = Pt LOSS
Pv1
(Pressure Loss Internal to System)
Figure 5N - Fan with Inlet Resistance
(AMCA Installation Type C)
Air Systems | 25
5.6 Types of fan system
PsF = Pt loss + PvO - Pv2
There are four basic system Installation Types:
In the special case where the velocity pressure at the
outlet is equal to the velocity pressure at the fan
discharge, the fan static pressure will equal the total
pressure loss. If these velocity pressures are different
the fan total pressure and fan static pressure must be
increased or decreased by the amount of the
difference in these velocity pressures.
AMCA INSTALLATION TYPE A:
Free Inlet, Free Outlet
5.6.3 AMCA Installation Type C: Fan system with
ducted inlet and free outlet. Figure 5N illustrates a
system with all system losses on the inlet side of the
fan. Since the velocity pressure at the outlet equals
the velocity pressure of the fan discharge and is also
equal to the total pressure at the fan discharge, the
fan static pressure will be equal to the total pressure
losses of the system.
AMCA INSTALLATION TYPE B:
Free Inlet, Ducted Outlet
PsF = Pt loss
AMCA INSTALLATION TYPE C:
Ducted Inlet, Free Outlet
Fans designed for use at the end of a system, such
as power roof ventilators, include the loss at the
discharge in the fan ratings and no system effect loss
is needed.
Figure 5P - Installation Types
5.6.1 AMCA Installation Type A: Fan system with
free inlet and free outlet. An AMCA Type A
installation covers equipment such as window fans,
panel fans and power roof ventilators. This type of
installation, in which there is no resistance at the fan
inlet or outlet, is shown in Figure 5M. The fan
provides the total pressure necessary to move air to
the velocity at the fan outlet. The total pressure of the
fan for this special case is equal to the fan velocity
pressure Pv2.
5.6.2 AMCA Installation Type B: Fan system with
free inlet and ducted outlet. Figure 5K shows a fan
system with discharge resistance. In this system:
Where:
PsO = Ps1 + Pv1 = 0
And:
26 | Air Systems
Eq. 5.6-3
The fan in Figure 5N has a short discharge duct
which is the way ducted fans are normally tested. If a
fan is used at the end of a system and is not
equipped with a 2 to 3 diameter length of duct, the
system effect loss at the discharge must be included
when determining the total pressure loss.
AMCA INSTALLATION TYPE D:
Ducted Inlet, Ducted Outlet
PtF = Pt loss + PvO
Eq. 5.6-2
Eq. 5.6-1
5.6.4 AMCA Installation Type D: Fan system with
ducted inlet and ducted outlet. Figure 5L illustrates
a system with system resistance on both inlet and
discharge sides of the fan. In this case the general
equations 5.5-3 and 5.5-4 apply. In the special case
where the velocity pressure at the outlet is equal to
the velocity pressure of the fan discharge, the fan
static pressure is equal to the total pressure loss
internal to the system.
The system shown in Figure 5L has higher velocity
pressure, relative to the static pressure, than would
normally be expected in a system. This is done to
emphasize the velocity pressure effects on system
total pressure losses for purposes of illustration.
To illustrate the effect of the outlet velocity on the total
pressure requirements, Figure 5Q shows the same
system as in Figure 5L with the addition of an evasé
outlet (diffuser) on the system. The evasé outlet
greatly reduces the system outlet velocity pressure.
The reduction in fan total pressure and fan static
pressure is clearly evident when compared to the
system in Figure 5L and is equal to the reduction in
the velocity pressure at the duct outlet minus the total
pressure loss in the evasé section. Losses will be
quite small for a long evasé outlet of good design
(optimum included angle is about 10 degrees). The
change in fan static pressure, because of the change
in the outlet velocity, is accounted for in the terms
"PvO - PvF" in Equation 5.5-4.
When system velocities exceed 15 m/s (3000 ft/min),
consideration should be given to the use of an evasé
outlet to reduce the system pressure requirements.
See ASHRAE Handbook, Fundamentals, chapter on
Duct Design for more information on the subject.
5.7 System resistance factors
The flow through any system is proportional to the
square root of the pressure causing the flow. This
relationship, which defines the flow versus pressure
characteristics of a particular system, is very useful in
fan and system designs. The flow rate in any system
was given earlier in equation 5.3-4B and can be
restated here as:
Q = 1.29(Ps)0.5Ao
Eq. 5.7-1 SI
Q = 4005(Ps)0.5Ao
Eq. 5.7-1 I-P
For standard air, where:
SR =
CoPvo
(Q / 1.29)2
Eq. 5.7-4 SI
SR =
CoPvo
(Q / 4005)2
Eq. 5.7-4 I-P
5.7.1 System resistance factors in series. The
particular value of using the system resistance factor
SR is that for resistances in series, the SR factors of
each element can be added to determine the system
resistance factor of the total system. As an example,
the three Figure 5R resistance factors in series can
be added to obtain the system resistance factor of the
complete system.
In this case, the resistance factors (given for SI and
(I-P) units, respectively) of 34.49 (0.3), 22.99 (0.2),
and 17.25 (0.15), total 74.73 (0.65). The pressure
loss of this system would be defined by the equation:
Pt loss = (Q/1.29)2SR
Eq. 5.7-5 SI
Pt loss = (Q/4005)2SR
Eq. 5.7-5 I-P
For an airflow rate of 2.83 m3/s (6000 cfm):
SI:
Ao = area of flow nozzle with no loss
Pt loss
This can also be stated as:
Pt = (Q/1.29)2SR
Eq. 5.7-2 SI
Pt = (Q/4005)2SR
Eq. 5.7-2 I-P
Where:
I-P:
Pt loss
SR = System resistance factor, m-4 (ft-4)
= 1/Ao2
The system resistance factor can be calculated from
known pressure loss information:
SR =
ΔPt
(Q / 1.29)2
Eq. 5.7-3 SI
SR =
ΔPt
(Q / 4005)2
Eq. 5.7-3 I-P
for standard air. They can also be calculated from the
dynamic loss coeffient:
= (Q/1.29)2SR
= (2.83/1.29)2 (34.49 + 22.99 + 17.25)
= (2.195)2 (74.73)
= (4.818)(74.73)
= 360 Pa
= (Q/4005)2SR
= (6000/4005)2 (0.3 + 0.2 + 0.15)
= (1.498)2 (0.65)
= (2.244)(0.65)
= 1.45 in. wg
5.7.2 System resistance factors in parallel. Similar
relationships can be established for flow through
parallel systems. The total pressure loss through
each branch of a parallel system must be equal to
establish equilibrium.
In Figure 5S the system resistance factor of each
branch is given as 0.3 and 0.2; and since the
pressure loss will be equal in both branches we can
equate these losses.
Air Systems | 27
RESISTANCE
PLANE 1
ELEMENT
PLANE 2
RESISTANCE
ELEMENT
E
0
ENTRY
FLOW
ATMOSPHERIC
PRESSURE
Pv
FAN
OUTLET
Pv2
PtF
Pt2
Pv0 = Pt0
Ps2
+
PtE = 0
_
0
Ps
PsF
Pt
ABSOLUTE
PRESSURE
Ps1
Pvt
Figure 5Q - Fan with Inlet and Discharge Resistance - Evasé Outlet Added
RESISTANCE 1
RESISTANCE 2
RESISTANCE 3
SR1 = 34.49 m-4
(0.3 ft.-4)
SR2 = 22.99 m-4
(0.2 ft.-4)
SR2 = 17.25 m-4
(0.15 ft.-4)
Figure 5R - Resistance in Series
RESISTANCE 1
Q1
SR1 = 34.49 m-4
(0.3 ft.-4)
QT
RESISTANCE 2
SR2 = 22.99 m-4
(0.2 ft.-4)
Figure 5S - Resistance in Parallel
28 | Air Systems
Q2
2
2
⎡ Q1 ⎤
⎡ Q2 ⎤
⎢ C ⎥ SR1 = ⎢ C ⎥ SR2
⎣ ⎦
⎣ ⎦
⎡ 0.4472 ⎤
= QT ⎢
⎥
⎣ 0.9949 ⎦
Where:
= QT (0.4495)
C = 1.29 (for SI units)
= 4005 (for I-P units)
If:
And:
QT = 0.472 m3/s or (1000 cfm)
SR2
Q1
=
Q2
SR1
And:
⎡ S ⎤
Q1 = Q2 ⎢ R2 ⎥
⎢⎣ SR1 ⎥⎦
Q1 = QT (0.4495)
Eq. 5.7-6
Q1 = (0.472 m3/s)(0.4495) or (1000 cfm)(0.4495)
= 0.212 m3/s or (449.5 cfm)
Substituting:
Q2 = QT - Q1
⎡
⎤
SR2
Q1 = QT ⎢
⎥
⎢⎣ SR1 + SR2 ⎥⎦
For the example shown in Figure 5S:
Then:
And:
Eq. 5.7-7
Q2 = QT - Q1
= 0.472 m3/s - 0.212 m3/s or (1000 cfm - 449.5 cfm)
= 0.260 m3/s or (550.5 cfm)
It can further be demonstrated that the relationship of
the system resistance factors for parallel systems is:
SI:
⎡
⎤
SR2
Q1 = QT ⎢
⎥
⎢⎣ SR1 + SR2 ⎥⎦
⎡
⎤
22.99
= QT ⎢
⎥
⎢⎣ 34.49 + 22.99 ⎥⎦
4.795
⎡
⎤
= QT ⎢
⎥
⎣ 5.873 + 4.795 ⎦
⎡ 4.795 ⎤
= QT ⎢
⎥
⎣10.667 ⎦
= QT ( 0.4495 )
I-P:
⎡
⎤
SR2
Q1 = QT ⎢
⎥
⎢⎣ SR1 + SR2 ⎥⎦
⎡
⎤
0 .2
= QT ⎢
⎥
⎢⎣ 0.3 + 0.2 ⎥⎦
⎡
⎤
0.4472
= QT ⎢
⎥
+
(
0
.
5477
0
.
4472
)
⎣
⎦
1
1
1
1
=
+
+ ... +
SRT
SR1
SR2
SRn
System resistance factors can be quite useful in
many system design and analysis problems.
5.8 System design and loss calculations
There are a number of design methods for sizing duct
work such as Equal Friction, Static Regain, etc.,
which are commonly used. The specific details of
these various methods will not be covered in this
publication, and the reader is referred to the
ASHRAE Guide and the Industrial Ventilation Guide
on this subject.
The procedure for calculating the total pressure loss
of the system is included as it is vital to the selection
of the fan.
Figure 5T, shown previously, will be used as an
example of the method of calculating the system total
pressure loss and fan static pressure required.
It is necessary to include in the loss calculation all
factors that contribute pressure loss, including
System Effects. The general method is to determine
the loss of each element of the system as they occur
Air Systems | 29
and total the losses. In systems that include parallel
branches it is customary to determine the loss of the
branch path that has the highest resistance to
establish the fan requirements. Other branch paths
that have lower losses must have resistance added
to them to balance the system since the total
pressure loss of each path must be equal.
5.8.1 Example: System loss calculations for
branch 1 (dynamic loss coefficient method)
ITEM K - DISCHARGE DIFFUSER
0.3048 m (12 in.) diameter with a dynamic loss
coefficient of 0.50.
(1000 cfm)
Airflow
= 0.47 m3/s
Pv
= 25.08 Pa
(0.101 in. wg)
Pt Loss
= 0.50 × 25.08
(0.50 × 0.101)
= 12.54 Pa
(0.051 in. wg)
NOTE: Loss coefficient includes velocity lost at
discharge and static regained by diffusion.
ITEM J - BALANCING DAMPER,
0.3048 m (12 in.) diameter with a dynamic loss
coefficient of 0.52.
(1000 cfm)
Airflow
= 0.47 m3/s
Pv
= 25.08 Pa
(0.101 in. wg)
Pt Loss
= 0.52 × 25.08
(0.52 × 0.101)
= 13.04 Pa
(0.053 in. wg)
ITEM DUCT,
6.096 m (20 ft.) of 0.3048 m (12 in.) diameter.
Airflow
= 0.47 m3/s
(1000 cfm)
= 10.18 Pa
(0.041 in. wg)
Pt Loss
(from Annex D)
ITEM A - ELBOW,
0.3048 m (12 in.) diameter with dynamic loss
coefficient of 0.22.
Airflow
= 0.47 m3/s
(1000 cfm)
Pv
= 25.08 Pa
(0.101 in. wg)
Pt Loss
= 0.22 × 25.08
(0.22 × 0.101)
= 5.518 Pa
(0.022 in. wg)
ITEM DUCT,
6.096 m (20 ft.) of 0.3048 m (12 in.) diameter.
Airflow
= 0.47 m3/s
(1000 cfm)
= 10.18 Pa
(0.041 in. wg)
Pt Loss
ITEM C - DIVIDED FLOW FITTING,
0.4064 m (16 in.) diameter to two 0.3048 m (12 in.)
diameter with a main branch dynamic loss coefficient
of 0.12.
Airflow
= 0.94 m3/s
(2000 cfm)
Pv
= 32.29 Pa
(0.13 in. wg)
Pt Loss
= 0.12 × 32.29
(0.12 × 0.13)
= 3.88 Pa
(0.016 in. wg)
30 | Air Systems
ITEM DUCT,
6.096 m (20 ft.) of 0.4064 m (16 in.) diameter.
(2000 cfm)
Airflow
= 0.94 m3/s
= 8.94 Pa
(0.036 in. wg)
Pt Loss
ITEM D - DIVIDED FLOW FITTING,
0.4572 m (18 in.) to 0.4064 m (16 in.) and 0.3048 m
(12 in.) diameter, with a main branch dynamic loss
coefficient of 0.
(3000 cfm)
Airflow
= 1.416 m3/s
Pv
= 44.70 Pa
(0.18 in. wg)
Pt Loss
= 44.70 × 0
(0.18 × 0)
= 0 Pa
(0 in. wg)
NOTE: The net loss in the main branch of this fitting
is zero (0), since there is static regain to offset the loss.
ITEM DUCT,
12.192 m (40 ft.), 0.4572 m (18 in.) diameter.
(3000 cfm)
Airflow
= 1.42 m3/s
Pt Loss
= 21.36 Pa
(0.086 in. wg)
E through H -These losses may be included in the
manufacturer's rating data.
ITEM E - ENTRANCE,
from plenum 0.4572 m (18 in.) diameter, dynamic
loss coefficient of 0.5.
(3000 cfm)
Airflow
= 1.42 m3/s
Pv
= 44.70 Pa
(0.18 in. wg)
Pt Loss
= 44.70 × 0.5
(0.18 × 0.5)
= 22.35 Pa
(0.09 in. wg)
ITEM F - COIL
Airflow
= 1.416 m3/s
Pt Loss
= 74.51 Pa
(from manufacturer's data)
(3000 cfm)
(0.3 in. wg)
ITEM G - FILTER
Airflow
= 1.416 m3/s
Pt Loss
= 86.93 Pa
(from manufacturer's data)
(3000 cfm)
(0.35 in. wg)
ITEM FAN,
Bulkhead Discharge (SEF due to lack of fan
discharge ductwork).
Airflow
= 1.416 m3/s
(3000 cfm)
Fan vel. = 6.53 m/s
(1285 fpm)
SEF
= 39.74 Pa
(0.16 in. wg)
(from AMCA Publication 201)
ITEM FAN,
Fan enclosed in a cabinet; SEF due to plenum wall
being too close to fan inlet).
(3000 cfm)
Airflow
= 1.42 m3/s
Inlet vel. = 5.31 ms
(1045 fpm)
SEF
= 7.45 Pa
(0.03 in. wg)
(from AMCA Publication 201)
ITEM H - INTAKE LOUVER
Airflow
= 1.42 m3/s
Pt Loss
= 19.87 Pa
(from manufacturer's data)
(3000 cfm)
(0.08 in. wg)
SYSTEM TOTAL PRESSURE LOSS BRANCH 1 =
336.4 Pa (1.355 in. wg).
5.8.2 Example: System loss calculations for
branch 2:
ITEM K - DIFFUSER
Pt Loss
= 12.42 Pa
(0.050 in. wg)
NOTE: Loss coefficient includes velocity lost at
discharge and static regained by diffusion.
ITEM J - DAMPER
Pt Loss
= 12.91 Pa
(0.052 in. wg)
ITEM DUCT,
6.096 m (20 ft.) of 0.3048 m (12 in.) diameter.
= 10.18 Pa
(0.041 in. wg)
Pt Loss
ITEM B - 45° ELBOW,
0.3048 m (12 in.) diameter, dynamic loss coefficient
of 0.13.
Airflow
= 0.47 m3/s
(1000 cfm)
Pv
= 25.08 Pa
(0.101 in. wg)
Since the elbow is located one duct diameter from
the divided flow fitting there will be a System Effect
loss that will need to be included.
K factor from Figure 4H = 1.08
Pt Loss
= 25.08 × 0.13 × 1.08 (0.101 × 0.13 × 1.08)
= 3.52 Pa
(0.014 in. wg)
ITEM C - DIVIDED FLOW FITTING,
0.4064 m (16 in.) diameter to 0.3048 m (12 in.)
diameter with a branch dynamic loss coefficient of
0.46.
Airflow
= 0.94 m3/s
(2000 cfm)
Pv
= 32.29 Pa
(0.13 in. wg)
K factor from Figure 4H = 1.08
= 0.46 × 32.29 × 1.08 (0.46 × 0.13 × 1.08)
Pt Loss
= 16.04 Pa
(0.065 in. wg)
Losses for the balance of the system are the same as
those previously calculated. These total 281.1 Pa
(1.132 in. wg).
SYSTEM TOTAL PRESSURE LOSS BRANCH 2 =
336.2 Pa (1.354 in. wg).
The loss through branch 2 is equal to the loss
through branch 1, and the branches are in balance.
Similar calculations for branch 3 show that it has a
lower resistance than branches 1 and 2, and some
resistance will need to be added by adjusting its
damper to balance the system.
5.8.3 Fan selection for the example system. The
total system pressure loss of the branch with the
highest resistance must be used to determine the fan
pressure requirements. The fan will need to provide a
total pressure of 336.2 Pa (1.354 in. wg) at 1.42 m3/s
(3000 cfm). The fan static pressure PsF is equal to the
fan total pressure PtF minus the calculated fan
discharge velocity pressure Pv2.
PsF = PtF - Pv2
= 336.2 - 25.58
= 310.62 Pa or (1.354 - 0.103 = 1.251 in. wg)
This procedure applies to any system regardless of
its complexity. The important point is that all losses,
including System Effect losses, need to be included
in the calculations. In this, as in many systems, the
System Effect losses are a significant portion of the
total pressure loss. AMCA Publication 201 should be
used to determine the System Effect losses for
various fan inlet and outlet conditions. Figure 5W
shows the point of operation of this system, where
the system resistance curve intersects the fan
performance curve.
5.9 Density effects in air systems
Since the density of the air varies with temperature,
pressure (altitude), and humidity, it is necessary to
evaluate the effect of density on the system design
and fan performance. Because of the variations in
density encountered in all air systems, a standard
density was established, and is used as the basis for
determining fan performance and system pressure
losses. Standard air density is defined as air with a
density of 1.2 kg/m3 (0.075 lbm/ft3). Fan performance
ratings and system element pressure loss ratings are
based on handling air at standard density.
The system designer must evaluate the actual air
density that will be handled by the system in order to
properly determine the volume of flow required and
the actual pressure losses in the system. Since fans
are essentially constant volume machines, the
volume of air handled by the fan will remain constant
regardless of the density, but the total pressure
developed by the fan and the power required by the
fan will vary in direct proportion to the density.
Similarly, the pressure losses in the system due to
friction and shock losses will also vary directly with
density.
In many applications it is actually the mass flow rate
that is important, and, therefore, the volume of air
required should be determined from the mass flow
Air Systems | 31
12.2m
6.1m
6.1m
FAN
D
1.42m3/s
0.46m DIA.
E
COIL
F
FILTER
G
LOUVER
H
C
A
0.94m3/s
0.472m3/s
0.41m DIA.
0.30m DIA.
B
B
0.30m DIA.
0.472m3/s
0.30m DIA.
0.472m3/s
J
J
DIFFUSER K
DAMPER
K
J
K
3
2
1
SI
40 ft.
20 ft.
20 ft.
FAN
D
3000 CFM
18” DIA.
E
COIL
F
FILTER
G
LOUVER
H
C
1000 CFM
16” DIA.
12” DIA.
B
B
12” DIA.
1000 CFM
12” DIA.
1000 CFM
J
J
DIFFUSER K
2
I-P
Figure 5T - Typical Air System
DAMPER
J
K
K
3
32 | Air Systems
A
2000 CFM
1
0.3810m DWDI FAN
0.2165M2 OUTLET AREA
1316 RPM
600
PRESSURE, Pa
DESIGN
SYSTEM
300
PT vs FLOW
0
0.5
1.0
1.5
2.0
3
m /sec
15” DWDI FAN
2.33 FT2 OUTLET AREA
1316 RPM
PRESSURE, H2O in. wg
2
DESIGN
SYSTEM
1
PT vs FLOW
0
1
2
3
4
CFM × 1000
Figure 5W - Fan Performance Versus System
Air Systems | 33
rate required at the design conditions. As an
example, in a system requiring a mass flow rate of
226.8 kg/min (500 lbm/min) at 121.1°C (250°F) and
at an altitude of 914.4 m (3000 ft), the air density from
Annex E is 0.800 kg/m3 (.05 lbm/ft3). The required
volume flow rate can then be determined by dividing
the mass flow rate by the design density:
conditions (Annex E) has a density of 1.072 kg/m3
(0.067 lbm/ft3). The power at these conditions would
be the power for standard air multiplied by the density
ratio:
SI (Volume flow rate)
226.8 kg/min ÷ (60 s/min × 0.800 kg/m3) = 4.725 m3/s
I-P (Power required at 68°F and 3000 ft)
14.5 BHP × (0.067 lbm/ft3 ÷ 0.075 lbm/ft3) = 12.95 BHP
I-P (Volume flow rate)
500 lbm/min ÷ 0.5 lbm/ft3 = 10000 cfm
In normal HVAC applications the effects of density
changes other than for operation at higher altitudes,
are quite often ignored and the system design is
based on handling standard air. The system designer
should, however, be aware of the effects of density
change and take them into consideration when
making field measurements of system performance
or balancing the system.
The pressure loss in the system would be calculated
based on 4.725 m3/s (10000 cfm) at standard air
density of 1.2 kg/m3 (.075 lbm/ft3). If in this example
the pressure loss at standard air is 1490.16 Pa (6 in.
wg) the pressure loss at actual conditions would be
this value multiplied by the density ratio:
SI (Pressure loss - actual conditions)
1490.15 Pa × (0.8 kg/m3 ÷ 1.2 kg/m3) = 993.44 Pa
I-P (Pressure loss - actual conditions)
6 in. wg × (.05 lbm/ft3 ÷ .075 lbm/ft3) = 4 in. wg
A fan for this system must be selected based on its
performance at standard conditions. For this example
select the fan for a performance of 4.725 m3/s (10000
cfm) at 1490.16 Pa (6 in. wg) static pressure.
Determine the fan power required at standard density
from the fan performance data. The power required
at actual conditions would then be calculated by
multiplying the catalog fan power by the density ratio.
In this example the fan power required at standard
conditions is 10.82 kW (14.5 BHP). At the actual
operating conditions of 121.1°C (250°F) air at 914.4
m (3000 ft) altitude, the fan power required would be
the power for standard air multiplied by the density
ratio:
SI (Power at actual conditions)
10.82 kW × (0.8 kg/m3 ÷ 1.2 kg/m3) = 7.21 kW
I-P (Power at actual conditions)
14.5 BHP × (0.05 lbm/ft3) ÷ .075 lbm/ft3) = 9.67 BHP
Consider other density effects when selecting a fan
for elevated temperature operations. If in this
example the fan would be required to start and run
some period of time at normal temperature as the
system warms up, the motor should be selected on
the basis of the cold air density. Since the fan will be
at 914.4 m (3000 ft) altitude, the air density at 20°C
(68°F) and 914.4 m (3000 ft) should be used to
determine the required fan power. Air at these
34 | Air Systems
SI (Power required at 20°C and 914.4 m)
10.82 kW × (1.072 kg/m3 ÷ 1.2 kg/m3) = 9.67 kW
Use Annex C to determine the density of air over a
range of barometric pressures, temperatures, and
relative humidities. Annex E gives the density ratios
for a wide range of temperatures and altitudes. The
air density at the various conditions is obtained by
multiplying standard air density of 1.2 kg/m3 (0.075
lbm/ft3) by the factors shown in the table.
6. System Design and Tolerances
Before making a final determination of the fan
selection there are several factors in the design of a
system and the selection of a fan that need to be
understood and evaluated:
a) The effect of variation in the resistance of the
actual installed system versus the resistance of
the designed system, i.e.; point of operation
b) The fan performance characteristics
system/performance tolerances
and
c) The effect of changes in the system, either
intentionally or unintentionally, on the point of
operation
d) The upper and lower system resistance design
points in systems that have variable resistance
characteristics (constant volume systems), or
that
have
variable
fan
performance
characteristics (variable air volume systems).
6.1 Point of operation
The system resistance of the actual installed system
can vary substantially from that calculated for the
system design, because of a number of factors:
a) The installed system is different from the
designed system, such as the addition of elbows
and offsets to meet field conditions, failure to
provide turning vanes in elbows, or the change in
position of various system elements with respect
to each other;
These curves are only shown to make the system
designer aware of the effect of the slope of the fan
curve on the expected system performance with
variations in system resistance. Many other factors
enter into the determination of the best fan type and
size for a given application.
b) Excessive leakage or increased resistance due
to poor quality workmanship at the installation;
6.2.2 Fan performance tolerance. The fan
performance also has a tolerance which must be
considered. The AMCA Check Test Tolerances are
described in AMCA Publication 211, Certified Ratings
Program -Air Performance, Product Rating
Requirement Subsection B. The AMCA Check Test
Tolerance is shown on Figure 6B. This tolerance is to
be applied along a parabolic system line. The power
required by the AMCA Check Test Tolerances shall
not exceed the rated data at the measured volume by
more than 5% or 37 watts, whichever is greater. The
fan curve in Figure 6B has dashed lines indicating the
tolerance range of fan performance, and when
combined with the system resistance tolerance
curves, an area of probable system performance is
indicated by the tolerance limits shown for the system
resistance and the tolerance limits shown for the fan
flow-pressure. As can be seen, the probable flow
range could be from 3.15 m3/s (6674 cfm) to 3.39
m3/s (7185 cfm) which is -4.7 to +2.6%. An installed
system tolerance range approaching ± 5.0% of flow
could be expected.
c) Loss coefficients of the various system elements
such as coils, filters, dampers, diffusers, elbows,
etc., improperly accounted for;
d) System Effects:
• not properly accounted for
• ignored in the original system design
• not accounted for because of on-site
installation changes.
The degree to which all of the various tolerances and
the field changes affect the actual system resistance
varies quite widely. Experience indicates that the
difference to be expected between the calculated and
actual system resistance can be as much as ±10%.
In extreme cases, greater system resistance
differentials have been experienced. Not accounting
for system effects in the design will result in a higher
system resistance and reduced flow.
6.2 Fan performance
There is a wide variety of basic fan designs in axial,
centrifugal and mixed flow variations. Curves of
several typical basic fans, all selected for a point of
operation of 3.30 m3/s (7000 cfm) at a static pressure
of 1490.16 Pa (6 in. wg), are shown on Figure 6A. All
of these fans pass through the design point of
operation, but with different slopes. The point of
highest efficiency will typically occur somewhat to the
right of the peak pressure point.
6.2.1 System resistance effect on performance.
The system resistance curves for the point of
operation and curves for ±10% and ±25% of the
design pressure are also shown in Figure 6A. The
intersection of these system lines with the various fan
curves show what range of volume performance can
be expected for each fan over this range of system
resistance. The amount of variation in flow rate with
changes in the system resistance will be dependent
on the slope of the fan performance curve in the
range of operation.
6.2.3 Performance safety factor. Evaluate the fan
performance tolerance and system resistance
tolerances to determine if the lower or upper limits of
the probable flow in the system are acceptable. The
combination of these tolerances should also be
evaluated to ensure that the high side system
resistance curve does not fall into the unstable
portion of the fan curve. With a few exceptions, all
fans have an unstable range of performance.
Operation in this area of the curve should be avoided
and precautions taken to ensure operation outside of
the unstable area at the highest expected system
resistance.
6.2.3.1 Static pressure safety factor. It has been
common practice among system designers to apply a
performance safety factor to the calculated system
requirements. This is often accomplished by adding a
nominal percentage of pressure to the system
pressure requirements. Some system designers will
size the system for a higher flow rate than is required.
The use of safety factors is discouraged when all
system components and system effects are
properly accounted for. The use of safety factors
is not required when system effect factors and all
known losses are accounted for.
Air Systems | 35
6.3 Effects of system changes
Some air systems are designed to operate at more
than one system condition, such as an exhaust
system serving multiple inlets where some of the
inlets can be closed off, or supply systems where
some of the outlets can be closed off or dampered for
reduced flow. The effect of these changes in the
system need to be evaluated in the system design
and the selection of a fan for this service. The main
concern would be that the fan is not forced to operate
in the unstable range. Also, the fan performance
should be such that the system performance is
acceptable over the range of operating conditions
desired. The motor must be selected to cover this
range of operating conditions. The system resistance
and system performance, for the example used in
Section 5.8, were calculated for the design condition
and also for two other system conditions. [Figure 6C
illustrates the effect of system changes.] The system
resistance and flow were calculated for the condition
where all the dampers were open and for the
condition where one of the dampers was closed.
Each of these conditions has a different system
resistance curve resulting in a different operating
point on the fan curve and a different total flow for the
system. The flow in each branch of the system will
also change. This leads to an important conclusion:
IN A FIXED SYSTEM, A CHANGE IN RESISTANCE
IN ANY ELEMENT WILL CHANGE THE TOTAL
SYSTEM RESISTANCE, AND AS A RESULT,
CHANGE THE POINT OF OPERATION ON THE
FAN CURVE AND THE FLOW RATE THROUGH ALL
OTHER ELEMENTS OF THE SYSTEM.
It is because of this interaction of the total system
with changes in any part of the system that the job of
balancing a system is very difficult.
6.4 Variable systems
Where systems are designed to be variable over
some range of operation, or where both the fan and
system are variable, the point of operation needs to
be evaluated at the upper and lower limits of
operation, relative to the tolerance of the fan and
system.
36 | Air Systems
Figure 6D shows a typical fan curve with system
resistance curves for a variable system where the
system pressure is allowed to vary as the system
demand for airflow changes. The system may be
varied by volume control, dampers, or other control
devices to provide a varying flow rate as demanded
by the system.
The tolerance ranges are shown for both the fan and
the system resistance. The most critical point in the
design of this type of system will be at the low flowhigh pressure condition. The fan selection and
system limits should be such that the fan will operate
in the stable portion of the fan curve at the maximum
resistance condition. Make sure the fan power
requirements over the tolerance range can be met by
the motor selected.
Figure 6E shows a typical set of fan curves for a
centrifugal fan with inlet vane control with system
resistance curves for a variable resistance system. In
this system, both the pressure and flow
characteristics of the fan are varied by changing the
inlet vanes position to meet the flow rate demand of
the system. Similar systems employ axial fans with
variable pitch control, or fans with variable speed
capability.
The critical area of fan selection is near the peak of
the pressure curve. Almost all fans exhibit some
degree of instability to the left of the peak pressure
point. It is wise to avoid operation in this range
without the expressed approval of the fan
manufacturer.
There are many system variations to meet various
design criteria that the designer may encounter. Not
all of the possibilities can be covered in the scope of
this publication. If fan users apply the principles
outlined in this publication to the specific system,
they can expect to design a good, functional system
and avoid many of the pitfalls often encountered in air
systems.
PRESSURE
1490.16 Pa
(6.0 in. wg)
+25%
+10%
-10%
BACKWARD
INCLINED FAN
-25%
RADIAL
BLADE FAN
VANEAXIAL
FAN
3.3m/s
(7000 cfm)
VOLUME FLOWRATE
Figure 6A - Performance of Example Fans with System Varations
+10%
DESIGN POINT
OF OPERATION
PRESSURE
-10%
FAN CURVE
POTENTIAL SYSTEM
FLOW RANGE
3.15 m3/s
(6701 cfm)
3.39 m3/s
(7210 cfm)
3.3 m3/s
(7000 cfm)
AMCA CERTIFIED RATING
TOLERANCE
VOLUME FLOWRATE
Figure 6B - Air Performance with Certified Ratings Tolerance
Air Systems | 37
DWDI FAN
TOTAL PRESSURE
DESIGN
SYSTEM
ONE DAMPER CLOSED
ALL DAMPERS CLOSED
Pt vs FLOW
VOLUME FLOWRATE
Figure 6C - Fan and System Curves - System Changes
+10%
HIGH PRESSURE
DESIGN POINT
-10%
FAN CURVE
PRESSURE
+10%
LOW PRESSURE
DESIGN POINT
AMCA CERTIFIED RATING
TOLERANCE
VOLUME FLOWRATE
Figure 6D - Fan and System Curves - Variable Resistance System
38 | Air Systems
-10%
PRESSURE
+10%
-10%
MAXIMUM FLOW
DESIGN POINT
VARIABLE VOLUME
SYSTEM RESISTANCE
IVC SETTING 1
IVC SETTING 2
IVC SETTING 3
IVC SETTING 4
STATIC PRESSURE
CONTROL POINT
VOLUME FLOWRATE
Figure 6E - Fan and System Curves - Variable Volume System
Air Systems | 39
Annex A. SI / I-P Conversion Table
Conversion factors between SI and I-P systems:
Quantity
I-P to SI
SI to I-P
Length
(ft) 0.3048 = m
(m) 3.2808 = ft
Mass (weight)
(lbs) 0.4536 = kg
(kg) 2.2046 = lbs.
Time
The unit of time is the second in both systems
Velocity
(ft-s) 0.3048 = ms
(ft/min) 0.00508 = ms
(ms) 3.2808 = fts
(ms) 196.85 = ft/min
Acceleration
(in./s2) 0.0254 = m/s2
(m/s2) 39.370 = in.s/2
Area
(ft2) 0.09290 = m2
(m2) 10.764 = ft2
Volume Flow Rate
(cfm) 0.000471948 = m3/s
(m3/s) 2118.88 = cfm
Density
(lb/ft3) 16.01846 = kg/m3
(kg/m3) 0.06243 = lb/ft3
Pressure
(in. wg) 248.36 = Pa
(in. wg) 0.24836 = kPa
(in. wg) 3.3864 = kPa
(Pa) 0.004026 = in. wg
(kPa) 4.0264 = in. wg
(kPa) 0.2953 = in. Hg
Viscosity:
Absolute
Kinematic
(lbm/ft-s) 1.4882 = Pa s
(ft2/s) 0.0929 = m2/s
(Pa s) 0.6719 = (lbm/ft-s)
(m2/s) 10.7639 = ft2/s
Gas Constant
(ft lb/lbm-°R) 5.3803 = J-kg/K
(j-kg/K) 0.1858 = (ft lb/lbm-°R)
Temperature
(°F - 32°)/1.8 = °C
(1.8 × °C) + 32° = °F
Power
(BHP) 746 = W
(BHP) 0.746 = kW
(W)/746 = BHP
(kW)/0.746 = BHP
40 | Air Systems
Annex B. Standard Atmospheric Data Versus Altitude Charts
Chart B.1 - SI Standard Atmospheric Data Versus Altitude
Z
Altitude
t
Temperature
p
Atmospheric
Pressure
ρ
Gas
Density
μ
Absolute
Viscosity
v
Kinematic
Viscosity
c
Speed of
Sound
m
°C
kPa
kg/m3
Pa•s
m2/s
m/s
0
100
200
300
15.00
14.35
13.70
13.05
101.32
100.13
98.94
97.77
1.230
1.215
1.201
1.189
1.793x10-5
1.790x10-5
1.786x10-5
1.784x10-5
1.456×10-5
1.473×10-5
1.487×10-5
1.500×10-5
340.43
340.05
339.66
339.28
400
500
600
700
12.40
11.76
11.11
10.46
96.61
95.46
94.32
93.20
1.177
1.166
1.155
1.145
1.780x10-5
1.777x10-5
1.774x10-5
1.771x10-5
1.512×10-5
1.524×10-5
1.536×10-5
1.546×10-5
338.89
338.51
338.19
337.73
800
900
1000
1100
9.81
9.16
8.51
7.86
92.08
90.98
89.88
88.80
1.134
1.123
1.112
1.102
1.768x10-5
1.765x10-5
1.761x10-5
1.758x10-5
1.559×10-5
1.571×10-5
1.584×10-5
1.595×10-5
337.34
336.95
336.57
336.18
1200
1300
1400
1500
7.21
6.56
5.90
5.25
87.72
86.66
85.61
84.56
1.091
1.080
1.069
1.058
1.755x10-5
1.751x10-5
1.748x10-5
1.745x10-5
1.609×10-5
1.621×10-5
1.635×10-5
1.649×10-5
335.79
335.40
335.01
334.62
1600
1700
1800
1900
4.60
3.95
3.30
2.65
83.53
82.50
81.49
80.49
1.047
1.037
1.026
1.016
1.741x10-5
1.738x10-5
1.735x10-5
1.732x10-5
1.663×10-5
1.676×10-5
1.691×10-5
1.705×10-5
334.22
333.83
333.44
333.05
2000
2100
2200
2300
2.00
1.35
0.70
0.53
79.49
78.51
77.54
76.57
1.006
0.996
0.986
0.976
1.728x10-5
1.725x10-5
1.722x10-5
1.718x10-5
1.718×10-5
1.732×10-5
1.746×10-5
1.760×10-5
332.66
332.26
331.87
331.48
2400
2500
2600
2700
-0.60
-1.25
-1.90
-2.55
75.62
74.68
73.74
72.82
0.967
0.957
0.948
0.938
1.715x10-5
1.712x10-5
1.708x10-5
1.705x10-5
1.774×10-5
1.789×10-5
1.802×10-5
1.818×10-5
331.08
330.69
330.29
329.90
2800
2900
3000
3100
-3.20
-3.85
-4.50
-5.15
71.91
71.00
70.11
69.23
0.929
0.919
0.909
0.900
1.702x10-5
1.699x10-5
1.695x10-5
1.692x10-5
1.832×10-5
1.845×10-5
1.865×10-5
1.880×10-5
329.50
329.10
328.71
328.31
3200
3300
3400
3500
-5.80
-6.46
-7.11
-7.76
68.35
67.48
66.62
65.77
0.890
0.880
0.871
0.862
1.689x10-5
1.685x10-5
1.682x10-5
1.679x10-5
1.898×10-5
1.914×10-5
1.931×10-5
1.948×10-5
327.51
327.11
326.70
326.70
Air Systems | 41
Chart B.2 - I-P Standard Atmospheric Data Versus Altitude
Z
Altitude
t
Temperature
p
Atmospheric
Pressure
ρ
Air
Density
μ
Dynamic
Viscosity
ν
Kinematic
Viscosity
c
Speed of
Sound
ft
°F
in. Hg
lbm/ft3
lbm/ft-s
ft2/s
ft/s
-1000
-500
62.6
60.8
31.02
30.47
.0787
.0776
1.212×10-5
1.208×10-5
1.538×10-4
1.556×10-4
1120.7
1118.8
0
59.0
29.92
.0765
1.205×10-5
1.576×10-4
1116.9
500
1000
1500
2000
57.2
55.4
53.7
51.9
29.38
28.86
28.33
27.82
.0754
.0743
.0732
.0721
1.202×10-5
1.198×10-5
1.195×10-5
1.192×10-5
1.595×10-4
1.614×10-4
1.633×10-4
1.653×10-4
1115.0
1113.1
1111.1
1109.2
2500
3000
3500
4000
50.1
48.3
46.5
44.7
27.32
26.82
26.33
25.84
.0710
.0700
.0689
.0679
1.189×10-5
1.185×10-5
1.182×10-5
1.179×10-5
1.673×10-4
1.694×10-4
1.714×10-4
1.735×10-4
1107.3
1105.3
1103.4
1101.4
4500
5000
5500
6000
43.0
41.2
39.4
37.6
25.37
24.90
24.43
23.98
.0669
.0659
.0649
.0639
1.175×10-5
1.172×10-5
1.169×10-5
1.165×10-5
1.757×10-4
1.778×10-4
1.800×10-4
1.823×10-4
1099.5
1097.5
1095.6
1093.6
6500
7000
7500
8000
35.8
34.0
32.3
30.5
23.53
23.09
22.65
22.22
.0630
.0620
.0610
.0601
1.162×10-5
1.158×10-5
1.155×10-5
1.152×10-5
1.846×10-4
1.869×10-4
1.892×10-4
1.916×10-4
1091.7
1089.7
1087.7
1085.7
8500
9000
9500
10000
28.7
26.9
25.1
23.3
21.80
21.39
20.98
20.58
.0592
.0583
.0574
.0565
1.148×10-5
1.145×10-5
1.142×10-5
1.138×10-5
1.904×10-4
1.965×10-4
1.990×10-4
2.015×10-4
1083.8
1081.8
1079.8
1077.8
42 | Air Systems
Annex C. Psychrometric Density Tables
Chart C.1 - Psychrometric Density Table (SI)
Density of Saturated Air for Various Barometric Conditions - kg/m3
Dry-Bulb
Temp. °C
-2.0
-1.5
-1.0
-0.5
0.0
0.5
1.0
1.5
2.0
2.5
3.0
3.5
4.0
4.5
5.0
5.5
6.0
6.5
7.0
7.5
8.0
8.5
9.0
9.5
10.0
10.5
11.0
11.5
12.0
12.5
13.0
13.5
14.0
14.5
15.0
15.5
16.0
16.5
17.0
17.5
18.0
Barometric Pressure kPa
97
1.244981
1.242122
1.239396
1.236782
1.234260
1.231812
1.229423
1.227079
1.224768
1.222480
1.220207
1.217942
1.215680
1.213416
1.211147
1.208871
1.206587
1.204295
1.201994
1.199687
1.197375
1.195060
1.192743
1.190428
1.188116
1.185810
1.183512
1.181224
1.178948
1.176683
1.174432
1.172192
1.169963
1.167742
1.165527
1.163312
1.161092
1.158860
1.156606
1.154320
1.151991
98.5
1.263273
1.260977
1.258667
1.256345
1.254012
1.251672
1.249325
1.246973
1.244618
1.242261
1.239902
1.237545
1.235188
1.232834
1.230483
1.228135
1.225792
1.223453
1.221119
1.218791
1.216468
1.214150
1.211838
1.209530
1.207227
1.204927
1.202631
1.200338
1.198047
1.195757
1.193466
1.191174
1.188879
1.186581
1.184277
1.181965
1.179644
1.177313
1.174968
1.172609
1.170232
100
1.282390
1.280094
1.277753
1.275377
1.272975
1.270553
1.268119
1.265679
1.263236
1.260796
1.258360
1.255931
1.253510
1.251098
1.248697
1.246304
1.243921
1.241546
1.239179
1.236817
1.234459
1.232105
1.229752
1.227399
1.225045
1.222689
1.220330
1.217968
1.215603
1.213236
1.210866
1.208497
1.206131
1.203771
1.201420
1.199084
1.196770
1.194483
1.192231
1.190025
1.187875
101.5
1.302927
1.300086
1.297353
1.294710
1.292141
1.289629
1.287163
1.284731
1.282324
1.279934
1.277553
1.275177
1.272800
1.270421
1.268037
1.265645
1.263247
1.260842
1.258431
1.256015
1.253595
1.251173
1.248752
1.246334
1.243920
1.241512
1.239113
1.236723
1.234343
1.231974
1.229616
1.227266
1.224925
1.222588
1.220251
1.217911
1.215560
1.213191
1.210795
1.208361
1.205877
103
1.324194
1.322000
1.319731
1.317400
1.315018
1.312595
1.310140
1.307661
1.305166
1.302659
1.300147
1.297634
1.295123
1.292618
1.290121
1.287634
1.285157
1.282692
1.280239
1.277798
1.275367
1.272946
1.270533
1.268128
1.265728
1.263332
1.260938
1.258544
1.256148
1.253747
1.251342
1.248928
1.246506
1.244075
1.241632
1.239178
1.236712
1.234235
1.231747
1.229250
1.226746
104.5
1.340401
1.337965
1.335505
1.333027
1.330532
1.328024
1.325506
1.322979
1.320447
1.317912
1.315376
1.312841
1.310307
1.307778
1.305254
1.302735
1.300224
1.297720
1.295225
1.292738
1.290260
1.287790
1.285328
1.282875
1.280428
1.277988
1.275553
1.273122
1.270693
1.268266
1.265837
1.263406
1.260970
1.258527
1.256073
1.253607
1.251125
1.248624
1.246101
1.243553
1.240975
Air Systems | 43
Chart C.1 - Psychrometric Density Table (SI) (Continued)
Density of Saturated Air for Various Barometric Conditions - kg/m3
Dry-Bulb
Temp. °C
18.5
19.0
19.5
20.0
20.5
21.0
21.5
22.0
22.5
23.0
23.5
24.0
24.5
25.0
25.5
26.0
26.5
27.0
27.5
28.0
28.5
29.0
29.5
30.0
30.5
31.0
31.5
32.0
32.5
33.0
33.5
34.0
34.5
35.0
35.5
36.0
36.5
37.0
37.5
38.0
44 | Air Systems
Barometric Pressure kPa
97
1.148567
1.146325
1.144073
1.141813
1.139548
1.137279
1.135008
1.132735
1.130461
1.128188
1.125917
1.123646
1.121378
1.119111
1.116846
1.114582
1.112318
1.110055
1.107790
1.105523
1.103253
1.100978
1.098695
1.096404
1.094102
1.091787
1.089456
1.087106
1.084735
1.082339
1.079915
1.077460
1.074970
1.072440
1.069867
1.067247
1.064575
1.061846
1.059056
1.056198
98.5
1.167391
1.164887
1.162437
1.160033
1.157668
1.155335
1.153029
1.150742
1.148470
1.146207
1.143949
1.141691
1.139431
1.137164
1.134888
1.132601
1.130299
1.127983
1.125650
1.123300
1.120932
1.118548
1.116147
1.113730
1.111299
1.108856
1.106402
1.103942
1.101478
1.099014
1.096553
1.094100
1.091661
1.089240
1.086844
1.084478
1.082149
1.079865
1.077632
1.075460
100
1.185062
1.182780
1.180492
1.178197
1.175897
1.173591
1.171279
1.168962
1.166639
1.164311
1.161977
1.159639
1.157295
1.154946
1.152592
1.150234
1.147871
1.145503
1.143131
1.140754
1.138373
1.135988
1.133599
1.131206
1.128809
1.126408
1.124004
1.121596
1.119184
1.116769
1.114351
1.111930
1.109506
1.107079
1.104649
1.102216
1.099780
1.097342
1.094902
1.092459
101.5
1.203323
1.200987
1.198647
1.196304
1.193957
1.191607
1.189254
1.186898
1.184537
1.182174
1.179806
1.177435
1.175060
1.172681
1.170298
1.167912
1.165521
1.163126
1.160726
1.158323
1.155915
1.153503
1.151086
1.148664
1.146239
1.143808
1.141372
1.138932
1.136487
1.134037
1.131582
1.129122
1.126656
1.124186
1.121710
1.119229
1.116742
1.114250
1.111753
1.109249
103
1.225071
1.222584
1.220116
1.217665
1.215229
1.212804
1.210388
1.207980
1.205577
1.203177
1.200778
1.198380
1.195979
1.193576
1.191169
1.188756
1.186338
1.183912
1.181480
1.179039
1.176591
1.174134
1.171669
1.169195
1.166714
1.164226
1.161731
1.159230
1.156724
1.154213
1.151700
1.149185
1.146669
1.144155
1.141644
1.139139
1.136640
1.134151
1.131673
1.129210
104.5
1.240138
1.237641
1.235154
1.232675
1.230205
1.227740
1.225283
1.222830
1.220383
1.217939
1.215499
1.213061
1.210625
1.208190
1.205755
1.203320
1.200883
1.198445
1.196003
1.193559
1.191110
1.188656
1.186196
1.183730
1.181257
1.178775
1.176286
1.173786
1.171277
1.168756
1.166224
1.163679
1.161121
1.158549
1.155963
1.153361
1.150743
1.148108
1.145455
1.142784
Chart C.2 - Psychrometric Density Table (I-P)
Density of Saturated Air for Various Barometric Conditions - lbm/ft3
Dry-Bulb
Temp. °F
30
31
32
33
34
35
36
37
38
39
40
41
42
43
44
45
46
47
48
49
50
51
52
53
54
55
56
57
58
59
60
61
62
63
64
Barometric Pressure in. Hg
28.5
.07703
.07687
.07671
.07654
.07638
.07621
.07605
.07589
.07573
.07557
.07541
.07525
.07509
.07493
.07477
.07461
.07445
.07429
.07413
.07397
.07381
.07366
.07350
.07334
.07318
.07302
.07287
.07271
.07255
.07240
.07224
.07208
.07193
.07177
.07161
29.0
.07839
.07822
.07806
.07789
.07772
.07756
.07739
.07723
.07706
.07690
.07674
.07657
.07641
.07625
.07609
.07592
.07576
.07560
.07544
.07528
.07512
.07496
.07479
.07464
.07447
.07431
.07415
.07399
.07383
.07367
.07352
.07336
.07320
.07304
.07288
29.5
.07974
.07957
.07940
.07924
.07907
.07890
.07873
.07856
.07840
.07823
.07806
.07790
.07773
.07757
.07740
.07724
.07707
.07691
.07674
.07658
.07642
.07625
.07609
.07593
.07576
.07560
.07544
.07528
.07512
.07495
.07479
.07463
.07447
.07430
.07414
30.0
.08111
.08093
.08075
.08058
.08041
.08024
.07807
.07990
.07973
.07956
.07939
.07922
.09705
.07889
.07872
.07855
.07838
.07822
.07805
.07788
.07772
.07755
.07739
.07722
.07706
.07689
.07673
.07656
.07640
.07623
.07607
.07590
.07574
.07557
.07541
30.5
.08245
.08228
.08210
.08193
.08175
.08158
.08141
.08123
.08106
.08089
.08072
.08055
.08038
.08021
.08004
.07986
.07970
.07953
.07936
.07919
.07902
.07885
.07868
.07852
.07835
.07818
.07801
.07784
.07768
.07751
.07734
.07718
.07701
.07684
.07668
31.0
.08380
.08363
.08345
.08327
.08310
.08292
.08274
.08257
.08239
.08222
.08205
.08187
.08170
.08153
.08135
.08118
.08101
.08084
.08066
.08049
.08032
.08015
.07998
.07981
.07964
.07947
.07930
.07913
.07896
.07879
.07862
.07845
.07828
.07811
.07794
Air Systems | 45
Chart C.2 - Psychrometric Density Table (I-P)
Density of Saturated Air for Various Barometric Conditions - lbm/ft3
Dry-Bulb
Temp. °F
65
66
67
68
69
70
71
72
73
74
75
76
77
78
79
80
81
82
83
84
85
86
87
88
89
90
91
92
93
94
95
96
97
98
99
100
46 | Air Systems
Barometric Pressure in. Hg
28.5
.07145
.07130
.07114
.07098
.07083
.07067
.07051
.07035
.07020
.07004
.06988
.06972
.06956
.06940
.06925
.06909
.06893
.06877
.06861
.06845
.06829
.06812
.06796
.06780
.06764
.06748
.06731
.06715
.06698
.06682
.06665
.06648
.06632
.06615
.06598
.06581
29.0
.07272
.07256
.07240
.07224
.07208
.07192
.07176
.07160
.07144
.07128
.07112
.07096
.07080
.07064
.07048
.07032
.07015
.07000
.06983
.06967
.06950
.06934
.06917
.06901
.06885
.06868
.06852
.06835
.06818
.06801
.06785
.06768
.06751
.06734
.06717
.06700
29.5
.07398
.07382
.07366
.07350
.07333
.07317
.07301
.07285
.07268
.07252
.07236
.07220
.07203
.07187
.07171
.07155
.07138
.07122
.07105
.07089
.07072
.07056
.07039
.07022
.07005
.06989
.06972
.06955
.06938
.06921
.06904
.06887
.06870
.06853
.06835
.06818
30.0
.07525
.07508
.07492
.07475
.07459
.07442
.07426
.07410
.07393
.07377
.07360
.07343
.07327
.07310
.07294
.07277
.07261
.07244
.07227
.07211
.07194
.07177
.07160
.07143
.07126
.07109
.07092
.07075
.07058
.07041
.07024
.07006
.06989
.06972
.06954
.06937
30.5
.07651
.07634
.07618
.07601
.07584
.07568
.07551
.07534
.07517
.07501
.07484
.07467
.07451
.07434
.07417
.07400
.07383
.07366
.07349
.07333
.07316
.07299
.07281
.07264
.07247
.07230
.07213
.07195
.07178
.07161
.07143
.07126
.07108
.01091
.07073
.07055
31.0
.07770
.07760
.07744
.07727
.07710
.07693
.07676
.07659
.07642
.07625
.07603
.07591
.07574
.07557
.07540
.07523
.07506
.07489
.07472
.07454
.07437
.07420
.07403
.07385
.07368
.07351
.07333
.07316
.07298
.07280
.07263
.07245
.07227
.07209
.07191
.07174
50
0.1
20
0.2
0.3
0.5
0.4
0.7
1
2
3
4
5
7
10
63
1.2
1.4
1.6
50
1.8
2.0
80
2.5
3
10
100
4
0
3.5
12
5
5
6
7
200
8
0
10
16
9
12
14
0
5
31
1,000
50
0
2,000
0
5,000
80
63
0
40
0
AIR QUANTITY, L/s at 1.20 kg/m3 (ε = 0.09 mm)
500
18
16
0
20
20
25
20
12
10
00
10,000
30
35
50
45
40
00
20,000
16
30
25
70
80
50,000
60
00
40
20
50
30
00
25
70
s
50
LO
CI
TY
m/
90
200,000
9
20
0.2
0.3
0.5
0.4
0.7
1
2
3
4
5
7
10
20
30
40
50
70
100
400,000
mm
R,
00
VE
IAM
CT
D
31
DU
ET
E
40
100
Annex D. Friction Charts
Chart D.1 - SI
FRICTION LOSS, Pa/m
Reprinted by permission of the American Society of Heating, Refrigerating and Air-Conditioning Engineers, Atlanta, Georgia, from the 1993
ASHRAE Handbook-Fundamentals.
Air Systems | 47
50
0.01
50
0.02
0.03
0.04
0.05
0.08
0.1
0.2
0.3
0.4
0.5
0.8
100
200
0
30
5
0
40
6
0
50
0
500
0
70
0
60
7
80
0
00
10
8
90
9
00
12
1,000
00
14
10
20
0
18 0
00
16
00
00
24
12
00
28
2,000
14
1
16
2
5,000
36
30
32
10,000
00
18
45
20,000
00
40
50,000
70
40
AIR QUANTITY, cfm at 0.075 lb/ft3 (ε = 0.0003 ft)
18
3
20
22
24
26
4
VE
LO
90
5
80
00
70
65 00
0
60 0
0
55 0
0
50 0
00
45
00
40
00
36
00
32
00
100,000
0
200,000
00
80
00
0
00
12
00
10
90
m
TY
, fp
CI
10
0
IAM
7
50
55
60
80
CT
D
DU
in.
R,
48 | Air Systems
ET
E
10
400,000
0.01
0.02
0.03
0.04
0.05
0.08
0.1
0.2
0.3
0.4
0.5
0.8
1
2
3
4
5
7
10
Chart D.2 - I-P
FRICTION LOSS, in. of water per 100 ft of duct
Reprinted by permission of the American Society of Heating, Refrigerating and Air-Conditioning Engineers, Atlanta, Georgia, from the 1993
ASHRAE Handbook-Fundamentals.
4
3
Annex E. Air Density Correction Factor Charts
Chart E.1 - Air Density Correction Factor (SI)
(Multiply Standard Air Density, 0.075 lbm/ft3 × the Factor to obtain Density at Condition B.)
Altitude, m
Sea Level
300
600
900
1200
Barometer, mm Hg
kPa
759.97
101.32
733.47
97.79
707.46
94.32
682.43
90.98
657.90
87.71
Air Temp. °C -20
0
20
40
60
80
100
120
140
160
180
200
220
240
260
280
300
320
340
1.22
1.12
1.03
.95
.93
.85
.79
.77
.73
.69
.65
.62
.57
.57
.56
.54
.52
.50
.49
1.12
1.04
.97
.91
.86
.80
.76
.72
.69
.65
.61
.57
.56
.55
.53
.52
.50
.48
.46
1.08
1.01
.94
.88
.83
.78
.73
.70
.67
.63
.59
.55
.54
.53
.51
.50
.48
.49
.48
1.04
.97
.90
.85
.80
.75
.70
.67
.64
.61
.57
.53
.52
.51
.49
.48
.46
.45
.43
1.00
.94
.87
.81
.77
.72
.68
.65
.62
.59
.55
.51
.50
.49
.47
.46
.44
.43
.41
Altitude, m
1500
1800
2100
2400
2700
Barometer, mm Hg
kPa
634.34
84.57
611.3
81.50
588.98
78.52
567.17
75.62
546.30
72.83
Air Temp. °C -20
0
20
40
60
80
100
120
140
160
180
200
220
240
260
280
300
320
340
.97
.90
.84
.78
.74
.69
.65
.62
.60
.57
.53
.49
.48
.47
.45
.44
.43
.41
.40
.93
.87
.81
.75
.71
.67
.63
.61
.58
.55
.51
.47
.46
.46
.44
.43
.41
.40
.39
.91
.84
.78
.73
.69
.65
.61
.59
.56
.53
.50
.47
.45
.46
.43
.42
.40
.39
.37
.87
.81
.75
.70
.67
.63
.59
.57
.54
.51
.49
.45
.43
.43
.41
.40
.38
.37
.36
.84
.80
.72
.68
.64
.60
.58
.58
.53
.49
.47
.43
.42
.41
.39
.38
.37
.35
.34
Air Systems | 49
Chart E.2 - Air Density Correction Factor (I-P)
(Multiply Standard Air Density, 0.075 lbm/ft3 × the Factor to obtain Density at Condition B.)
Altitude, ft.
-1000
Sea Level
1000
2000
3000
4000
Barometer, in. Hg
in wg.
31.02
422.2
29.92
407.5
28.86
392.81
27.82
378.6
26.82
365.0
25.84
351.7
Air Temp. °F -40
0
40
70
100
150
200
250
300
350
400
450
500
550
600
700
800
900
1000
1.31
1.19
1.10
1.04
0.98
0.90
0.83
0.77
0.72
0.68
0.64
0.60
0.57
0.54
0.52
0.47
0.44
0.40
0.37
1.26
1.15
1.06
1.00
0.95
0.87
0.80
0.75
0.70
0.65
0.62
0.58
0.55
0.53
0.50
0.46
0.42
0.39
0.36
1.22
1.11
1.02
0.96
0.92
0.84
0.77
0.72
0.67
0.62
0.60
0.56
0.53
0.51
0.48
0.44
0.40
0.37
0.35
1.17
1.07
0.99
0.93
0.88
0.81
0.74
0.70
0.65
0.60
0.57
0.54
0.51
0.49
0.46
0.43
0.39
0.36
0.33
1.13
1.03
0.95
0.89
0.85
0.78
0.71
0.67
0.62
0.58
0.55
0.52
0.49
0.47
0.45
0.41
0.37
0.35
0.32
1.09
0.99
0.92
0.86
0.81
0.75
0.69
0.64
0.60
0.56
0.53
0.50
0.47
0.45
0.43
0.39
0.36
0.33
0.31
Altitude, ft.
5000
6000
7000
8000
9000
10,000
Barometer, in. Hg
in wg.
24.90
338.9
23.98
326.4
23.09
314.3
22.22
302.1
21.39
291.1
20.58
280.1
Air Temp. °F -40
0
40
70
100
150
200
250
300
350
400
450
500
550
600
700
800
900
1000
1.05
0.95
0.88
0.83
0.78
0.72
0.66
0.62
0.58
0.54
0.51
0.48
0.45
0.44
0.41
0.38
0.35
0.32
0.30
1.01
0.91
0.85
0.80
0.75
0.69
0.64
0.60
0.56
0.52
0.49
0.46
0.44
0.42
0.40
0.37
0.33
0.31
0.29
0.97
0.89
0.82
0.77
0.73
0.67
0.62
0.58
0.54
0.51
0.48
0.45
0.43
0.41
0.39
0.35
0.32
0.30
0.27
0.93
0.85
0.79
0.74
0.70
0.65
0.60
0.56
0.52
0.49
0.46
0.43
0.41
0.39
0.37
0.34
0.31
0.29
0.26
0.90
0.82
0.76
0.71
0.68
0.62
0.57
0.58
0.50
0.47
0.44
0.42
0.39
0.38
0.35
0.33
0.30
0.28
0.26
0.87
0.79
0.73
0.69
0.65
0.60
0.55
0.51
0.48
0.45
0.42
0.40
0.38
0.36
0.34
0.32
0.29
0.27
0.25
50 | Air Systems
Fans and Systems
ANSI/AMCA 210 Laboratory Methods of Testing Fans
For Aerodynamic Performance Rating, offers the
system design engineer guidance as to how the fan
was tested and rated. AMCA Publication 201 Fans
and Systems, helps provide guidance as to what
effect the system and its connections to the fan have
on fan performance.
Recognizing and accounting for losses that affect the
fan’s performance, in the design stage, will allow the
designer to predict with reasonable accuracy, the
installed performance of the fan.
1.1 Purpose
This part of the AMCA Fan Application Manual
includes general information about how fans are
tested in the laboratory, and how their performance
ratings are calculated and published. It also reviews
some of the more important reasons for the "loss" of
fan performance that may occur when the fan is
installed in an actual system.
Allowances, called System Effect Factors (SEF), are
also given in this part of the manual. SEF must be
taken into account by the system design engineer if a
reasonable estimate of fan/system performance is to
be determined.
1.2 Some limitations
It must be appreciated that the System Effect Factors
given in this manual are intended as guidelines and
are, in general, approximations. Some have been
obtained from research studies, others have been
published previously by individual fan manufacturers,
and many represent the consensus of engineers with
considerable experience in the application of fans.
Fans of different types and even fans of the same
type, but supplied by different manufacturers, will not
necessarily react with the system in exactly the same
way. It will be necessary, therefore, to apply judgment
based on actual experience in applying the SEF.
The SEF represented in this manual assume that the
fan application is generally consistent with the
method of testing and rating by the manufacturer.
Inappropriate application of the fan will result in SEF
values inconsistent with the values presented.
Mechanical design of the fan is not within the scope
of this publication.
201
2. Symbols and Subscripts
For symbols and subscripted symbols, see Table 2.1.
For subscripts, see Table 2.2.
3. Fan Testing
Fans are tested in setups that simulate installations.
The four standard installation types are as shown in
Figure 3.1.
AMCA INSTALLATION TYPE A:
Free Inlet, Free Outlet
AMCA INSTALLATION TYPE B:
Free Inlet, Ducted Outlet
AMCA INSTALLATION TYPE C:
Ducted Inlet, Free Outlet
AMCA INSTALLATION TYPE D:
Ducted Inlet, Ducted Outlet
Figure 3.1 - Standard Fan Installation Types
3.1 ANSI/AMCA Standard 210
Most fan manufacturers rate the performance of their
products from tests made in accordance with
ANSI/AMCA 210 Laboratory Methods of Testing Fans
for Aerodynamic Performance Rating. The purpose
of ANSI/AMCA 210 is to establish uniform methods
for laboratory testing of fans and other air moving
devices to determine performance in terms of airflow,
Table 2.1 - Symbols and Subscripted Symbols
SYMBOL
DESCRIPTION
UNITS OF MEASURE
SI
I-P
A
Area of cross section
m2
ft2
D
Diameter, impeller
mm
in.
D
Diameter, Duct
m
ft
H
Fan Power Input
kw
hp
H/T
Hub-to-Tip Ratio
Dimensionless
Kp
Compressibility Coefficient
Dimensionless
Cp
Loss Coefficient
Dimensionless
N
Speed of Rotation
rpm
rpm
Ps
Fan Static Pressure
Pa
in. wg
Pt
Fan Total Pressure
Pa
in. wg
Pv
Fan Velocity Pressure
Pa
in. wg
pb
Corrected Barometric Pressure
kPa
in. Hg
PL
Plane of Measurement
---
---
Q
Airflow
m3/s
ft3/min
Re
Fan Reynolds Number
SEF
System Effect Factor
Pa
in. wg
td
Dry-Bulb Temperature
°C
°F
tw
Wet-Bulb Temperature
°C
°F
μ
Air Viscosity
Pa•s
lbm/ft•s
V
Velocity
m/s
fpm
W
Power Input to Motor
watts
watts
ηs
Fan Static Efficiency
%
%
ηt
Fan Total Efficiency
%
%
ρ
Air Density
kg/m3
lbm/ft3
Dimensionless
Table 2.2 - Subscripts
SUBSCRIPT
DESCRIPTION
a
c
x
1
2
3
5
6
8
Atmospheric conditions
Converted Value
Plane 0, 1, 2, ...as appropriate
Fan Inlet Plane
Fan Outlet Plane
Pitot Traverse Plane
Plane 5 (nozzle inlet station in chamber)
Plane 6 (nozzle discharge station in chamber)
Plane 8 (inlet chamber measurement station)
52 | Fans and Systems
pressure, power, air density, speed of rotation and
efficiency, for rating or guarantee purposes. Two
methods of measuring airflow are included: the Pitot
tube and the long radius flow nozzle. These are
incorporated into a number of "setups" or "figures".
In general, a fan is tested on the setup that most
closely resembles the way in which it will be installed
in an air system. Centrifugal and axial fans are
usually tested with an outlet duct. Propeller fans are
normally tested in the wall of a chamber or plenum.
Power roof ventilators (PRV) are tested mounted on
a curb exhausting from the test chamber.
It is very important to realize that each setup in
ANSI/AMCA 210 is a standardized arrangement that
is not intended to reproduce exactly any installation
likely to be found in the field. The infinite variety of
possible arrangements of actual air systems makes it
impractical to duplicate every configuration in the fan
test laboratory.
3.2 Ducted outlet fan tests
Figure 3.2 is a reproduction of a test setup from
ANSI/AMCA 210. Note that this particular setup
includes a long straight duct connected to the outlet
of the fan. A straightener is located upstream of the
Pitot traverse to remove swirl and rotational
components from the airflow and to ensure that
airflow at the plane of measurement is as nearly
uniform as possible.
the fan outlet is limited to ensure that uniform airflow
will be maintained. A steep transition, or abrupt
change of cross section would cause turbulence and
eddies. The effect of this type of airflow disturbance
at the fan outlet is discussed later.
Uniform airflow conditions ensure consistency and
reproducibility of test results and permit the fan to
develop its maximum performance. In any installation
where uniform airflow conditions do not exist, the
fan's performance will be measurably reduced.
As illustrated in Figure 3.3 Plane 2, the velocity
profile at the outlet of a fan is not uniform. The section
of straight duct attached to the fan outlet controls the
diffusion of the outlet airflow and establishes a more
uniform velocity as shown in Figure 3.3 Plane X.
The energy loss when a gas, such as air, passes
through a sudden enlargement is related to the
square of the velocity. Thus the ducted outlet with its
more uniform velocity significantly reduces the loss at
the point of discharge to the atmosphere.
A manufacturer may test a fan with or without an inlet
duct or outlet duct. For products licensed to use the
AMCA Certified Ratings Seal, catalog ratings will
state whether ducts were used during the rating tests.
If the fans are not to be applied with the same duct(s)
as in the test setup, an allowance should be made for
the difference in performance that may result.
The angle of the transition between the test duct and
1
Transition
Piece
2
Straightener
FOR FAN INSTALLATION TYPES:
B: Free Inlet, Ducted Outlet D: Ducted Inlet, Ducted Outlet
Figure 3.2 - Pitot Traverse in Outlet Duct
Fans and Systems | 53
3.3 Free inlet, free outlet fan tests
Figure 3.4 illustrates a typical multi-nozzle chamber
test setup from ANSI/AMCA 210. This simulates the
conditions under which most exhaust fans are tested
and rated. Fan performance based on this type of
test may require adjustment when additional
accessories are used with the fan. Fans designed for
use without duct systems are usually rated over a
lower range of pressures. They are commonly
cataloged and sold as a complete unit with suitable
drive and motor.
3.4 Obstructed inlets and outlets
The test setups in ANSI/AMCA 210 result in
unobstructed airflow conditions at both the inlet and
the outlet of the fan. Appurtenances or obstructions
located close to the inlet and/or outlet will affect fan
performance. Shafts, bearings, bearing supports and
other appurtenances normally used with a fan should
be in place when a fan is tested for rating.
Variations in construction which may affect fan
performance include changes in sizes and types of
sheaves and pulleys, bearing supports, bearings and
shafts, belt guards, inlet and outlet dampers, inlet
vanes, inlet elbows, inlet and outlet cones, and
cabinets or housings.
Since changes in performance will be different for
various product designs, it will be necessary to make
suitable allowances based on data obtained from the
applicable fan catalog or directly from the
manufacturer.
Most single width centrifugal fans are tested using
Arrangement 1 fans. Some allowance for the effect
of bearings and bearing supports in the inlet may be
necessary when using Arrangement 3 or
Arrangement 7. The various AMCA standard
arrangements are shown on Figures 3.5, 3.6, and
3.7.
4. Fan Ratings
4.1 The Fan Laws
It is not practical to test a fan at every speed at which
it may be applied. Nor is it possible to simulate every
inlet density that may be encountered. Fortunately,
by use of a series of equations commonly referred to
as the Fan Laws, it is possible to predict with good
accuracy the performance of a fan at other speeds
and densities than those of the original rating test.
The performance of a complete series of
geometrically similar (homologous) fans can also be
54 | Fans and Systems
calculated from the performance of smaller fans in
the series using the appropriate equations.
Because of the relationship between the airflow,
pressure and power for any given fan, each set of
equations for changes in speed, size or density,
applies only to the same Point of Rating, and all the
equations in the set must be used to define the
converted condition. A Point of Rating is the specified
fan operating point on its characteristic curve.
The Fan Law equations are shown below as ratios.
The un-subscripted variable is used to designate the
initial or test fan values for the variable and the
subscript c is used to designate the converted,
dependent or desired variable.
Qc = Q × (Dc/D)3 × (Nc/N) × (Kp/Kpc)
Ptc = Pt × (Dc/D)2 × (Nc/N)2 × (ρc/ρ) × (Kp/Kpc)
Pvc = Pv × (Dc/D)2 × (Nc/N)2 × (ρc/ρ)
Psc = Ptc - Pvc
Hc = H × (Dc/D)5 × (Nc/N)3 × (ρc/ρ) × (Kp/Kpc)
ηtc = (Qc × Ptc × Kp) / Hc
(SI)
ηtc = (Qc × Ptc × Kp) / (6362 • Hc)
(I-P)
ηsc = ηtc × (Psc/Ptc)
These equations have their origin in the classical
theories of fluid mechanics, and the accuracy of the
results obtained is sufficient for most applications.
Better accuracy would require consideration of
Reynolds number, Mach number, kinematic viscosity,
dynamic viscosity, surface roughness, impeller blade
thickness and relative clearances, etc.
4.2 Limitations
Under certain conditions the properties of gases
change and there are, therefore, limitations to the use
of the Fan Laws. Accurate results will be obtained
when the following limitations are observed:
a. Fan Reynolds Number (Re). The term Reynolds
number is associated with the ratio of inertia to
viscous forces. When related to fans, investigations
of both axial and centrifugal fans show that
performance losses are more significant at low
Reynolds number ranges and are effectively
negligible above certain threshold Reynolds
numbers. In an effort to simplify the comparison of
the Reynolds numbers of two fans, the fan industry
PL X
BLAST AREA
PL 2
DISCHARGE DUCT
OUTLET AREA
CUTOFF
CENTRIFUGAL FAN
PL 2
PL X
AXIAL FAN
Figure 3.3 - Controlled Diffusion and Establishment of a Uniform Velocity
Profile in a Straight Length of Outlet Duct
38mm ±6mm
(1.5in. ±0.25 in.)
PL.5 PL.6
PL.8
PL.1
PL.2
0.5 M MIN.
0.2M
MIN.
0.5M
MIN.
0.2 M MIN.
0.3 M MIN.
t d2
AIRFLOW
M
FAN
VARIABLE
SUPPLY
SYSTEM
t d3
0.1 M MIN.
SETTLING
MEANS
SETTLING
MEANS
(See note 4)
Ps5
ΔP
Pt8
Figure 3.4 - Inlet Chamber Setup - Multiple Nozzles in Chamber
(ANSI/AMCA 210-99, Figure 15)
Fans and Systems | 55
ANSI/AMCA Standard 99-2404-03
Page 1 of 2
Drive Arrangements for Centrifugal Fans
An American National Standard - Approved by ANSI on April 17, 2003
AMCA Drive
Arrangement
1 SWSI
ISO 13349
Drive
Arrangement
1 or
12 (Arr. 1 with
sub-base)
Description
Fan Configuration
Alternative Fan
Configuration
For belt or direct drive.
Impeller overhung on shaft, two
bearings mounted on pedestal
base.
Alternative: Bearings mounted
on independant pedestals, with
or without inlet box.
2 SWSI
For belt or direct drive.
2
Impeller overhung on shaft,
bearings mounted in bracket
supported by the fan casing.
Alternative: With inlet box.
3 SWSI
3 or
11 (Arr. 3 with
sub-base)
For belt or direct drive.
Impeller mounted on shaft
between bearings supported by
the fan casing.
Alternative: Bearings mounted
on independent pedestals, with
or without inlet box.
3 DWDI
6 or
18 (Arr. 6 with
sub-base)
For belt or direct drive.
Impeller mounted on shaft
between bearings supported by
the fan casing.
Alternative: Bearings mounted
on independent pedestals, with
or without inlet boxes.
4 SWSI
4
For direct drive.
Impeller overhung on motor
shaft. No bearings on fan.
Motor mounted on base.
Alternative: With inlet box.
5 SWSI
5
For direct drive.
Impeller overhung on motor
shaft. No bearings on fan.
Motor flange mounted to
casing.
Alternative: With inlet box.
AMCA International, Inc. | 30 W. University Dr. | Arlington Heights, IL, 60004-1893 | U.S.A
Figure 3.5 - AMCA Standard 99-2404 / Page 1
56 | Fans and Systems
AMCA 201-02
Page 2 of 2
ANSI/AMCA Standard 99-2404-03
AMCA Drive
Arrangement
ISO 13349
Drive
Arrangement
7 SWSI
7
Description
Fan Configuration
Alternative Fan
Configuration
For coupling drive.
Generally the same as Arr. 3,
with base for the prime mover.
Alternative: Bearings mounted
on independent pedestals with
or without inlet box.
7DWDI
For coupling drive.
17
(Arr. 6 with
Generally the same as Arr. 3
base for motor) with base for the prime mover.
Alternative: Bearings mounted
on independent pedestals with
or without inlet box.
8 SWSI
8
For direct drive.
Generally the same as Arr. 1
with base for the prime mover.
Alternative: Bearings mounted
on independent pedestals with
or without inlet box.
9 SWSI
9
For belt drive.
Impeller overhung on shaft, two
bearings mounted on pedestal
base.
Motor mounted on the outside
of the bearing base.
Alternative: With inlet box.
10 SWSI
10
For belt drive.
Generally the same as Arr. 9
with motor mounted inside of
the bearing pedestal.
Alternative: With inlet box.
AMCA International, Inc. | 30 W. University Dr. | Arlington Heights, IL, 60004-1893 | U.S.A
Figure 3.6 - AMCA Standard 99-2404 / Page 2
Fans and Systems | 57
ANSI/AMCA Standard 99-3404-03
Page 1 of 1
Drive Arrangements for Axial Fans
An American National Standard - Approved by ANSI on June 10, 2003
Note: All fan orientations may be horizontal or vertical
AMCA Drive
Arrangement
1
ISO 13349
Drive
Arrangement
1
12 (Arr. 1 with
sub-base)
Description
Alternative Fan
Configuration
Fan Configuration
For belt or direct drive.
Impeller overhung on shaft, two
bearings mounted either
upstream or downstream of the
impeller.
Alternative: Single stage or two
stage fans can be supplied with
inlet box and/or discharge
evasé.
3
3
11 (Arr. 3 with
sub-base)
For belt or direct drive.
Impeller mounted on shaft
between bearings on internal
supports.
Alternative: Fan can be
supplied with inlet box, and/or
discharge evasé.
4
4
For direct drive.
Impeller overhung on motor
shaft. No bearings on fan.
Motor mounted on base or
integrally mounted.
M
M
M
M
Alternative: With inlet box
and/or with discharge evasé.
7
7
For direct drive.
Generally the same as Arr. 3
with base for the prime mover.
M
M
M
M
Alternative: With inlet box
and/or discharge evasé.
8
8
For direct drive.
Generally the same as Arr. 1
with base for the prime mover.
Alternative: Single stage or two
stage fans can be supplied with
inlet box and/or discharge
evasé.
9
9
For belt drive.
Generally same as Arr. 1 with
motor mounted on fan casing,
and/or an integral base.
Alternative: With inlet box
and/or discharge evasé
M
AMCA International, Inc. | 30 W. University Dr. | Arlington Heights, IL, 60004-1893 | U.S.A
Figure 3.7 - AMCA Standard 99-3404 / Page 1
58 | Fans and Systems
calculated using the proper specific heat ratio for the
gases being handled.
has adopted the term Fan Reynolds Number.
Re = (πND2ρ) / (60μ)
where: N = impeller rotational speed, rpm
D = impeller diameter, m(ft)
ρ = air density, kg/m3 (lbm/ft3)
μ = absolute viscosity,
1.8185 × 10-5 Pa•s (5°C to 38°C)
(1.22 × 10-5 lbm/ft•s (40°F to 100°F))
(SI)
(I-P)
The threshold fan Reynolds number for centrifugal
and axial fans is about 3.0 × 106. That is, there is a
negligible change in performance between the two
fans due to differences in Reynolds number if both
fans are operating above this threshold value. When
the Reynolds number of a model fan is below 3.0 ×
106, there may be a gain in efficiency (size effect) for
a full size fan operating above the threshold
compared to one operating below the threshold. This
occurs only when both fans are operating near peak
efficiency. Therefore, when a model test is being
conducted to verify the rating of a full size fan, the
Reynolds number should be above 3.0 ×106 to avoid
any uncertainty relating to Reynolds number effects.
b. Point of Rating. To predict the performance of a
fan from a smaller model using the Fan Laws, both
fans must be geometrically similar (homologous),
and both fans must operate at the same
corresponding rating points on their characteristic
curves. Two or more fans are said to be operating at
corresponding “points of rating” if the positions of the
operating points, relative to the pressure at shutoff
and the airflow at free delivery, are the same.
c. Compressibility. Compressibility is the characteristic
of a gas to change its volume as a function of
pressure, temperature and composition. The
compressibility coefficient (Kp) expresses the ratio of
the fan total pressure developed with an
incompressible fluid to the fan total pressure
developed with a compressible fluid (See
ANSI/AMCA 210). Differences in the compressibility
coefficient between two similar fans must be
d. Specific Heat Ratio (Cp). Model fan tests are
usually based on air with a specific heat ratio of 1.4.
Induced draft fans may handle flue gas with a specific
heat ratio of 1.35. Even though these differences may
normally be considered small, they make a
noticeable difference in the calculation of the
compressibility coefficient. Refer to AMCA
Publication 802, Annex A, for calculation procedures.
e. Tip Speed Mach Parameter (Mt). Tip speed Mach
parameter is an expression relating the tip speed of
the impeller to the speed of sound at the fan inlet
condition.
When airflow velocity at a point approaches the
speed of sound, some blocking or choking effects
occur that reduce the fan performance.
4.3 Fan performance curves
A fan performance curve is a graphic presentation of
the performance of a fan. Usually it covers the entire
range from free delivery (no obstruction to airflow) to
no delivery (an air tight system with no air flowing).
One, or more, of the following characteristics may be
plotted against volume airflow (Q).
Fan Static Pressure
Fan Total Pressure
Fan Power
Fan Static Efficiency
Fan Total Efficiency
Ps
Pt
H
ηs
ηt
Air density (ρ), fan size (D), and fan rotational speed
(N) are usually constant for the entire curve and must
be stated.
A typical fan performance curve is shown in Figure
4.1. Figure 4.2 illustrates examples of performance
curves for a variety of fan types.
Fans and Systems | 59
SIZE 30 FAN AT N RPM
Pt
100
Ps
ηt
ηs
70
60
50
40
H
30
20
OPERATION AT
STANDARD DENSITY
10
0
AIRFLOW, Q
Figure 4.1 - Fan Performance Curve at N RPM
60 | Fans and Systems
EFFICIENCY, η PERCENT
80
POWER, H
PRESSURE, P
90
TYPE
BACKWARDINCLINED
BACKWARDCURVED
HOUSING DESIGN
• Highest efficiency of all centrifugal fan designs.
• Ten to 16 blades of airfoil contour curved away from
direction of rotation. Deep blades allow for efficient
expansion within blade passages
• Air leaves impeller at velocity less than tip speed.
• For given duty, has highest speed of centrifugal
fan designs
• Scroll-type design for efficient conversion of velocity
pressure to static pressure.
• Maximum efficiency requires close clearance and
alignment between wheel and inlet
• Efficiency only slightly less than airfoil fan.
• Ten to 16 single-thickness blades curved or inclined
away from direction of rotation
• Efficient for same reasons as airfoil fan.
• Uses same housing configuration as airfoil design.
• Higher pressure characteristics than airfoil,
backward-curved, and backward-inclined fans.
• Curve may have a break to left of peak pressure
R
and fan should not be operated in this area.
• Power rises continually to free delivery.
RADIAL
CENTRIFUGAL FANS
AIRFOIL
IMPELLER DESIGN
R
• Scroll. Usually narrowest of all centrifugal designs.
• Because wheel design is less efficient, housing
dimensions are not as critical as for airfoil and
backward-inclined fans.
M
CENTRIFUGAL
AXIAL
CENTRIFUGAL
A
POWER ROOF VENTILATORS
SPECIAL DESIGNS
TUBULAR
VANEAXIAL
TUBEAXIAL
AXIAL FANS
PROPELLER
FORWARDCURVED
M
B
• Flatter pressure curve and lower efficiency than the
airfoil, backward-curved, and backward-inclined.
• Do not rate fan in the pressure curve dip to the left
of peak pressure.
• Power rises continually toward free delivery. Motor
selection must take this into account.
• Scroll similar to and often identical to other centrifugal
fan designs.
• Fit between wheel and inlet not as critical as for airfoil
and backward-inclined fans.
• Low efficiency.
• Limited to low-pressure applications.
• Usually low cost impellers have two or more blades of
single thickness attached to relatively small hub.
• Primary energy transfer by velocity pressure.
• Simple circular ring, orifice plate, or venturi.
• Optimum design is close to blade tips and forms
smooth airfoil into wheel.
• Somewhat more efficient and capable of developing
more useful static pressure than propeller fan.
• Usually has 4 to 8 blades with airfoil or singlethickness cross section.
• Hub usually less than transfer by velocity pressure.
• Cylindrical tube with close clearance to blade tips.
• Good blade design gives medium- to high-pressure
capability at good efficiency.
• Most efficient of these fans have airfoil blades.
• Blades may have fixed, adjustable, or controllable
pitch.
• Hub is usually greater than half fan tip diameter.
• Cylindrical tube with close clearance to blade tips.
• Guide vanes upstream or downstream from impeller
increase pressure capability and efficiency.
• Performance similar to backward-curved fan except
capacity and pressure are lower.
• Lower efficiency than backward-curved fan.
• Performance curve may have a dip to the left of
peak pressure.
• Cylindrical tube similar to vaneaxial fan, except
clearance to wheel is not as close.
• Air discharges radially from wheel and turns 90° to
flow through guide vanes.
• Low-pressure exhaust systems such as general
factory, kitchen, warehouse, and some commercial
installations.
• Provides positive exhaust ventilation, which is an
advantage over gravity-type exhaust units.
• Centrifugal units are slightly quieter than
axial units.
• Normal housing not used, since air discharges from
impeller in full circle.
• Usually does not include configuration to recover
velocity pressure component.
• Low-pressure exhaust systems such as general
factory, kitchen, warehouse, and some commercial
installations.
• Provides positive exhaust ventilation, which is an
advantage over gravity-type exhaust units.
• Essentially a propeller fan mounted in a supporting
structure
• Hood protects fan from weather and acts as safety
guard.
• Air discharges from annular space at bottom of
weather hood.
Figure 4.2 - Types of Fans
Adapted with permission from 1996 ASHRAE Systems and Equipment Handbook (SI)
Fans and Systems | 61
PERFORMANCE CURVES
a
PERFORMANCE CHARACTERISTICS
Pt
8
Ps
10
6
ηt
4
8
6
ηs
wo
2
4
2
VOLUME FLOW RATE, Q
0
0
2
EFFICIENCY
PRESSURE-POWER
10
4
6
8
6
8
6
4
4
2
VOLUME FLOW RATE, Q
0
2
4
6
8
10
6
8
6
4
4
2
2
VOLUME FLOW RATE, Q
0
2
4
6
8
EFFICIENCY
PRESSURE-POWER
8
0
10
6
8
6
4
4
2
2
VOLUME FLOW RATE, Q
0
2
4
6
8
EFFICIENCY
PRESSURE-POWER
8
0
0
10
8
10
6
8
6
4
4
2
2
VOLUME FLOW RATE, Q
0
0
2
4
6
8
EFFICIENCY
PRESSURE-POWER
10
8
10
6
8
6
4
4
2
2
VOLUME FLOW RATE, Q
0
0
2
4
6
8
EFFICIENCY
PRESSURE-POWER
• Primarily for materials handling in industrial plants.
Also for some high-pressure industrial requirements.
• Rugged wheel is simple to repair in the field. Wheel
sometimes coated with special material.
• Not common for HVAC applications.
• Pressure curve less steep than that of backward-curved
fans. Curve dips to left of peak pressure.
• Highest efficiency to right of peak pressure at 40 to 50%
of wide open volume.
• Rate fan to right of peak pressure.
• Account for power curve, which rises continually toward
free delivery, when selecting motor.
• Primarily for low-pressure HVAC applications, such as
residential furnaces, central station units, and
packaged air conditioners.
• High flow rate, but very low-pressure capabilities.
• Maximum efficiency reached near free delivery.
• Discharge pattern circular and airstream swirls.
• For low-pressure, high-volume air moving applications,
such as air circulation in a space or ventilation through
a wall without ductwork.
• Used for makeup air applications.
• High flow rate, medium-pressure capabilities.
• Performance curve dips to left of peak pressure. Avoid
operating fan in this region.
• Discharge pattern circular and airstream rotates or swirls.
• Low- and medium-pressure ducted HVAC applications
where air distribution downstream is not critical.
• Used in some industrial applications, such as drying
ovens, paint spray booths, and fume exhausts.
• High-pressure characteristics with medium-volume flow
capabilities.
• Performance curve dips to left of peak pressure due to
aerodynamic stall. Avoid operating fan in this region.
• Guide vanes correct circular motion imprated by wheel and
improve pressure characteristics and efficiency of fan.
• General HVAC systems in low-, medium-, and high-pressure
applications where straight-through flow and compact
installation are required.
• Has good downstream air distribution
• Used in industrial applications in place of tubeaxial fans.
• More compact than centrifugal fans for same duty.
• Performance similar to backward-curved fan, except
capacity and pressure is lower.
• Lower efficiency than backward-curved fan because air
turns 90°.
• Performance curve of some designs is similar to axial flow
fan and dips to left of peak pressure.
• Primarily for low-pressure, return air systems in HVAC
applications.
• Has straight-through flow.
• Usually operated without ductwork; therefore, operates at
very low pressure and high volume.
• Only static pressure and static efficiency are shown for
this fan.
• Low-pressure exhaust systems, such as general factory,
kitchen, warehouse, and some commercial installations.
• Low first cost and low operating cost give an advantage over
gravity flow exhaust systems.
• Centrifugal units are somewhat quieter than axial flow units.
• Usually operated without ductwork; therefore, operates at
very low pressure and high volume.
• Only static pressure and static efficiency are shown for
this fan.
• Low-pressure exhaust systems, such as general factory,
kitchen, warehouse, and some commercial installations.
• Low first cost and low operating cost give an advantage
over gravity flow exhaust systems.
0
10
10
0
10
10
8
10
6
8
6
4
4
2
2
VOLUME FLOW RATE, Q
0
0
2
4
6
8
EFFICIENCY
PRESSURE-POWER
• Higher pressure characteristics than airfoil and backwardcurved fans.
• Pressure may drop suddenly at left of peak pressure, but
this usually causes no problems.
• Power rises continually to free delivery.
0
10
10
0
10
10
8
10
6
8
6
4
4
2
2
VOLUME FLOW RATE, Q
0
0
2
4
6
8
EFFICIENCY
PRESSURE-POWER
• Same heating, ventilating, and air-conditioning
applications as airfoil fan.
• Used in some industrial applications where airfoil
blade may corrode or erode due to environment.
0
10
10
0
10
10
8
10
6
8
6
4
4
2
2
VOLUME FLOW RATE, Q
0
0
2
4
6
8
EFFICIENCY
PRESSURE-POWER
• Similar to airfoil fan, except peak efficiency slightly lower.
10
EFFICIENCY
PRESSURE-POWER
8
0
0
10
10
8
10
6
8
6
4
4
2
2
0
0
VOLUME FLOW RATE, Q
2
4
6
8
EFFICIENCY
PRESSURE-POWER
• General heating, ventilating, and air-conditioning
applications.
• Usually only applied to large systems, which may be
low-, medium-, or high-pressure applications.
• Applied to large, clean-air industrial operations for
significant energy savings.
0
10
10
2
APPLICATIONS
• Highest efficiencies occur at 50 to 60% of wide open
volume. This volume also has good pressure
characteristics.
• Power reaches maximum near peak efficiency and
becomes lower, or self-limiting, toward free delivery.
0
10
a: These performance curves reflect general characteristics of various fans as commonly applied. They are not intended to provide complete selection criteria, since other parameters,
such as diameter and speed, are not defined.
Figure 4.2 - Types of Fans
Adapted with permission from 1996 ASHRAE Systems and Equipment Handbook (SI)
62 | Fans and Systems
5.1 Type A: Free inlet, free outlet fans
1) Type B: Free inlet, ducted outlet
2) Type C: Ducted inlet, free outlet
3) Type D: Ducted inlet, ducted outlet
Fans designed for use other than with duct systems
are usually rated over a lower range of pressures.
They are commonly cataloged and sold as a
complete unit with suitable drive and motor.
The performance of fans intended for use with duct
systems is usually published in the form of a "multirating" table. A typical multi-rating table, as illustrated
in Figure 5.2 shows:
Typical fans in this group are propeller fans and
power roof ventilators. They are usually available in
direct or belt-drive arrangements and performance
ratings are published in a modified form of the multirating table. Figure 5.1 illustrates such a table for part
of a line of belt-drive propeller fans.
a) the speed (N) in rpm
b) the power (H) in kw (hp)
c) the fan static pressure (Ps) in Pa (in. wg)
d) the outlet velocity (V) in m/s, (fpm)
e) the airflow (Q) in m3/s (cfm)
5. Catalog Performance Tables
5.2 Ducted fans
There are three types of ducted fans, as described in
Section 3:
Figure 5.3 shows constant speed characteristic
curves superimposed on a section of the multi-rating
table for the same fan. A brief study of this figure will
assist in understanding the relationship between
curves and the multi-rating tables.
SIZE No. of Motor
Peak
AIRFLOW (m3/s) @ STATIC PRESSURE (Pa)
rpm
(cm) Blades kW
kW
0
31
62
93
124
155
186
217
0.19
862
0.13
2.02
1.58
0.58
0.19
960
0.20
2.25
1.87
0.97
61
3
0.25 1071 0.27
2.51
2.18
1.76
0.76
0.37 1220 0.40
2.86
2.57
2.24
1.70
0.81
0.19
806
0.20
2.89
2.36
1.05
0.25
883
0.27
3.17
2.68
1.94
0.76
69
3
0.37 1035 0.43
3.71
3.30
2.85
1.56
0.95
0.56 1165 0.62
4.18
3.83
3.44
3.01
1.60
1.10
0.37
825
0.42
4.36
3.76
3.04
1.27
0.56
945
0.62
4.99
4.48
3.92
2.38
1.42
0.75
1045
0.82
5.23
5.08
4.57
4.01
2.31
1.52
84
3
1.12 1190 1.19
6.29
5.90
5.47
5.01
4.48
2.79
1.94
1.49 1306 1.64
6.91
6.53
6.15
5.75
5.32
4.81
3.05
2.24
TYPICAL RATING TABLE FOR A SERIES OF BELT-DRIVEN PROPELLER FANS
248
1.84
SIZE No. of Motor
Peak
AIRFLOW (ft3/min) @ STATIC PRESSURE (in. wg)
rpm
(in.) Blades hp
bhp
0
1/8
1/4
3/8
1/2
5/8
3/4
7/8
1
1/4
862
0.18 4,283 3,350 1,230
1/4
960
0.27 4,770 3,960 2,050
24
3
1/3
1071 0.36 5,321 4,620 3,730 1,600
1/2
1220 0.54 6,062 5,450 4,750 3,600 1,710
1/4
806
0.27 6,123 4,990 2,230
1/3
883
0.36 6,708 5,675 4,100 1,620
27
3
1/2
1035 0.57 7,862 7,000 6,035 3,315 2,020
3/4
1165 0.83 8,850 8,110 7,290 6,385 3,400 2,330
1/2
825
0.56 9,240 7,970 6,430 2,700
3/4
945
0.83 10,580 9,500 8,300 5,040 3,010
1
1045
1.1 11,710 10,755 9,685 8,490 4,890 3,215
33
3
1½
1190
1.6 13,335 12,490 11,580 10,610 9,500 5,905 4,100
2
1306
2.2 14,630 13,845 13,030 12,185 11,280 10,200 6,470 4,740 3,900
TYPICAL RATING TABLE FOR A SERIES OF BELT-DRIVEN PROPELLER FANS
Figure 5.1 - Propeller Fan Performance Table
Fans and Systems | 63
IMPELLER DIAMETER:
TIP SPEED IN m/s:
Volume
m3/s
1.81
2.17
2.53
2.89
3.25
3.61
3.97
4.33
4.69
5.06
5.42
5.78
6.14
6.50
6.86
7.22
7.94
8.67
9.39
10.11
10.83
11.55
12.28
13.00
13.72
14.44
Outlet
Vel.
(m/s)
2.55
3.06
3.56
4.07
4.58
5.08
5.59
6.10
6.61
7.13
7.63
8.14
8.65
9.15
9.66
10.17
11.18
12.21
13.23
14.24
15.25
16.27
17.30
18.31
19.32
20.34
62 Pa
rpm
kW
222
236
253
272
292
314
338
361
385
409
434
458
483
508
0.14
0.17
0.22
0.27
0.34
0.42
0.51
0.62
0.74
0.88
1.03
1.21
1.41
1.63
927 mm
.0485 × RPM
93 Pa
rpm
kW
270
284
300
317
337
358
379
402
426
449
473
498
522
547
571
621
0.25
0.30
0.36
0.43
0.52
0.62
0.74
0.87
1.01
1.18
1.37
1.58
1.81
2.06
2.34
2.99
OUTLET AREA:
MAXIMUM kW:
124 Pa
rpm
kW
313
327
343
360
378
398
419
441
464
488
511
535
559
585
633
682
0.39
0.45
0.53
0.63
0.73
0.86
1.00
1.16
1.33
1.53
1.75
1.99
2.25
2.54
3.20
3.98
155 Pa
rpm
kW
352
366
382
399
417
437
457
479
501
525
538
571
595
644
693
742
791
0.55
0.64
0.74
0.85
0.98
1.13
1.30
1.49
1.69
1.92
2.16
2.44
2.74
3.41
4.20
5.13
6.20
.71 SQ METERS
13.65 × (RPM/1000)3
186 Pa
217 Pa
rpm
kW
rpm
kW
389
403
419
436
454
473
494
515
537
560
584
607
654
703
752
801
850
0.75
0.86
0.98
1.11
1.26
1.44
1.63
1.86
2.09
2.35
2.62
2.93
3.63
4.44
5.38
6.47
7.70
411
424
438
455
472
489
509
529
550
572
595
616
665
712
761
810
859
908
0.87
0.98
1.10
1.25
1.41
1.58
1.79
2.01
2.26
2.54
2.82
3.14
3.85
4.68
5.64
6.73
7.99
9.40
246 Pa
rpm
kW
310 Pa
rpm
kW
373 Pa
rpm
kW
443 1.10
458 1.19 494 1.52
472 1.39 507 1.68 540 1.99
489 1.56 522 1.86 554 2.18
506 1.74 538 2.06 568 2.39
524 1.95 555 2.28 584 2.62
543 2.18 572 2.53 600 2.89
564 2.43 590 2.78 617 3.16
585 2.71 610 3.07 635 3.45
606 3.01 630 3.39 654 3.78
629 3.34 651 3.74 674 4.15
675 4.07 695 4.52 715 4.96
721 4.93 741 5.40 759 5.89
769 5.90 788 6.41 805 6.94
818 7.01 834 7.57 852 8.11
867 8.27 883 8.87 898 9.47
916 9.70 932 10.32 946 10.96
965 11.30 981 11.95 995 12.62
1015 13.06 1030 13.77 1044 14.46
1072 15.78 1093 16.50
1129 17.98 1142 18.76
TYPICAL MULTISPEED RATING TABLE FOR A SINGLE WIDTH, SINGLE INLET CENTRIFUGAL FAN
IMPELLER DIAMETER:
TIP SPEED IN FPM:
Volume
CFM
3825
4590
5355
6120
6885
7650
8415
9180
9945
10710
11475
12240
13005
13770
14535
15300
16830
18360
19890
21420
22950
24480
26010
27540
29070
30600
Outlet
Vel.
(fpm)
500
600
700
800
900
1000
1100
1200
1300
1400
1500
1600
1700
1800
1900
2000
2200
2400
2600
2800
3000
3200
3400
3600
3800
4000
36.5 IN
9.56 × RPM
OUTLET AREA:
MAXIMUM BHP:
7.65 SQ FT
18.3 × (RPM/1000)3
1/4 in. wg
3/8 in. wg
1/2 in. wg
5/8 in. wg
3/4 in. wg
7/8 in. wg
1 in. wg
1-1/4 in. wg 1-1/2 in. wg
rpm
bhp
rpm
bhp
rpm
bhp
rpm
bhp
rpm
rpm
rpm
rpm
222
236
253
272
292
314
338
361
385
409
434
458
483
508
0.185
0.233
0.292
0.365
0.450
0.560
0.682
0.826
0.989
1.175
1.387
1.626
1.895
2.191
270
284
300
317
337
358
379
402
425
449
473
498
522
547
571
621
0.334
0.400
0.483
0.579
0.695
0.832
0.988
1.163
1.360
1.587
1.837
2.115
2.424
2.767
3.144
4.003
313
327
343
360
378
398
419
441
464
488
511
535
559
585
633
682
0.519
0.608
0.716
0.840
0.981
1.149
1.340
1.553
1.780
2.048
2.346
2.665
3.017
3.403
4.289
5.335
352
366
383
399
417
437
457
479
501
525
538
571
595
644
693
742
791
0.743
0.856
0.992
1.144
1.314
1.514
1.741
1.993
2.269
2.570
2.901
3.275
3.672
4.577
5.632
6.885
8.308
bhp
389 1.01
403 1.15
419 1.31
436 1.49
454 1.69
473 1.93
494 2.19
515 2.49
537 2.80
560 3.15
584 3.52
607 3.93
654 4.87
703 5.96
752 7.22
801 8.67
850 10.32
bhp
bhp
bhp
rpm
411 1.17
424 1.31 443 1.48
438 1.48 458 1.60 494 2.04
455 1.68 472 1.86 507 2.25 540 2.67
472 1.89 489 2.09 522 2.49 554 2.92
489 2.12 506 2.34 538 2.76 568 3.20
509 2.40 524 2.61 555 3.06 584 3.52
529 2.70 543 2.92 572 3.39 600 3.87
550 3.03 564 3.26 590 3.73 617 4.24
572 3.40 585 3.64 610 4.12 635 4.63
595 3.78 606 4.04 630 4.55 654 5.07
618 4.21 629 4.48 651 5.02 674 5.56
665 5.16 675 5.46 695 6.06 715 6.65
712 6.28 721 6.61 741 7.24 759 7.90
761 7.56 769 7.91 788 8.60 805 9.30
810 9.03 818 9.40 834 10.15 852 10.88
859 10.71 867 11.09 883 11.89 898 12.70
908 12.50 916 13.01 932 13.84 946 14.70
965 15.16 981 16.03 995 16.92
1015 17.52 1030 18.47 1044 19.39
1072 21.16 1093 22.13
1129 24.11 1142 25.16
TYPICAL MULTISPEED RATING TABLE FOR A SINGLE WIDTH, SINGLE INLET CENTRIFUGAL FAN
Figure 5.2 - Centrifugal Fan Performance Tables
64 | Fans and Systems
bhp
OUTLET
VELOCITY
500
600
700
800
900
1000
1100
1200
1300
1400
1500
1600
1700
1800
1900
2000
2200
2400
2600
2800
3000
3200
3400
3600
3800
4000
VOLUME
CFM
3825
4590
5355
6120
6885
7650
8415
9180
9945
10710
11475
12240
13005
13770
14535
15300
16830
18360
19890
21420
22950
24480
26010
27540
29070
30600
.185
.233
.292
.365
.450
BHP
434 1.387
456 1.626
482
508 2.19
.334
.400
.483
.579
BHP
571 3.744
629 4.003
449 1.587
473 1.837
493 2.115
522 2.424
547 2.767
BHP
584 3.403
633 4.289
682 5.335
464 1.78
488 2.048
511 2.346
535 2.665
559 3.017
360 .840
378 .981
398 1.149
419 1.340
441 1.553
313 .519
327 .608
343 .716
RPM
1/2” SP
BHP
596
644 4.577
693 5.632
742 6.885
791 8.308
479 1.995
501 2.269
525 2.570
538 2.901
571 3.276
332 .992
399 1.144
417 1.314
437 1.514
457 1.741
352 .743
366 .856
RPM
5/8” SP
RPM
3.93
4.87
5.76
7.22
8.67
2.19
2.49
2.80
3.15
3.52
1.15
1.31
1.49
1.69
1.93
850 10.32
607
654
703
752
801
494
515
537
560
584
403
419
436
454
473
389 1.01
BHP
3/4” SP
390 RPM
337 .695
358 .822
379 .988
482 1.163
426 1.360
270
284
300
317
RPM
3/8” SP
BHP
4.21
5.16
6.28
7.56
9.03
2.40
2.70
3.03
3.40
1.31
1.48
1.58
1.89
2.12
859 10.71
908 12.60
618
665
712
761
810
509
529
550
572
595
424
438
455
472
489
411 1.17
RPM
7/8” SP
BHP
4.48
5.46
6.81
7.91
8.48
2.61
2.92
3.26
3.84
4.04
1.48
1.60
1.86
2.09
2.34
3.06
3.49
3.73
4.12
4.55
2.04
2.25
2.49
2.76
BHP
651 5.02
695 6.06
741 7.24
788 8.60
834 10.15
555
572
590
610
630
494
507
522
538
RPM
1-1/4” SP
BHP
3.52
3.87
4.24
4.63
5.07
674 5.56
715 6.65
759 7.90
9.30
852 10.88
584
600
617
635
654
540 2.67
554 2.92
568 3.28
RPM
1-1/2” SP
BHP
3.99
4.36
4.76
5.18
5.63
696 6.11
736 7.24
778
822 10.02
867 11.65
612
627
643
661
678
584 3.37
598 3.66
RPM
1-3/4” SP
867 11.09 883 11.89 898 12.70 914 13.48
916 13.01 932 13.84 946 14.70 960 15.56
965 15.16 981 16.03 995 16.92 1009 17.83
1015 17.52 1030 18.47 1044 19.39 1057 20.35
1079 21.16 1093 22.13 1106 23.12
1129 24.11 1142 25.16 1155 26.18
629
675
721
769
818
524
543
564
585
606
443
458
472
489
506
RPM
1” SP
490 RPM
314 .560
338 .682
361 .826
335 .988
409 1.175
222
236
253
272
292
RPM
1/4” SP
PRESSURE IN IN. WG
BRAKE HORSEPOWER
585 RPM
810 RPM
RECOMMENDED
SELECTION RANGE
CFM
Figure 5.3 - Typical Fan Performance Table Showing Relationship to a Family
of Constant Speed Performance Curves
Fans and Systems | 65
Most performance tables do not cover the complete
range from no delivery to free delivery but cover only
the typical operating range. Figure 5.4 illustrates the
recommended performance range of a centrifugal
fan. Comparison of Figure 5.4 with Figure 5.3 will
show that the published performance table also
covers only the recommended performance range of
the fan.
6. Air Systems
6.1 The system
An air system may consist simply of a fan with
ducting connected to either the inlet or outlet or to
both. A more complicated system may include a fan,
ductwork, air control dampers, cooling coils, heating
coils, filters, diffusers, sound attenuation, turning
vanes, etc. See AMCA Publication 200 Air Systems,
for more information.
It should be remembered that fans are generally
tested without obstructions in the inlet and outlet and
without any optional airstream accessories in place.
Catalog ratings will, therefore, usually apply only to
the bare fan with unobstructed inlet and outlet.
6.2 Component losses
Fan performance adjustment factors for airstream
accessories are normally available from either the fan
catalog or the fan manufacturer.
Every system has a combined resistance to airflow
that is usually different from every other system and
is dependent upon the individual components in the
system.
Fans are usually tested in arrangement 1, or similar
(see Figure 3.5). Rating tables will, therefore, also
apply
only
to
the
tested
arrangement.
Allowances for the effect of bearing supports used in
other arrangements should be obtained from the
manufacturer if not shown in the catalog.
The determination of the "pressure loss" or
"resistance to airflow," for the individual components
can be obtained from the component manufacturers.
The determination of pressure losses for ductwork
design is well documented in standard handbooks
such as the ASHRAE Handbook of Fundamentals.
SELECTION NOT USUALLY
RECOMMENDED IN THIS RANGE
RECOMMENDED
SELECTION RANGE
SY
ST
EM
CU
RV
E
PRESSURE
RE
SU
ES
PR
DU
CT
E
RV
EM
SELECTION
NOT USUALLY
RECOMMENDED
IN THIS RANGE
CU
ST
T
UC
SY
D
AIRFLOW
Figure 5.4 - Recommended Performance Range of a Typical Centrifugal Fan
66 | Fans and Systems
In a later section, the effects of some system
components and fan accessories on fan performance
are discussed. The System Effects presented will
assist the system designer to determine fan
selection.
The system curve of a "fixed system" plots as a
parabola in accordance with the above relationship.
Typical plots of the resistance to flow versus volume
airflow for three different and arbitrary fixed systems,
(A, B, and C) are illustrated in Figure 6.1. For a fixed
system an increase or decrease in airflow results in
an increase or decrease in the system resistance
along
the
given
system
curve
only.
Also, as the components in a system change, the
system curve changes.
6.3 The system curve
At a fixed airflow through a given air system a
corresponding pressure loss, or resistance to this
airflow, will exist. If the airflow is changed, the
resulting pressure loss, or resistance to airflow, will
also change. The relationship between airflow
pressure and loss can vary as a function of type of
duct components, their interaction and the local
velocity magnitude. In many cases, typical duct
systems operate in the turbulent flow regime and the
pressure loss can be approximated as a function of
velocity (or airflow) squared. The simplifying
relationship used in this publication governing the
change in pressure loss as a function of airflow for a
fixed system is:
Refer to Figure 6.1, Duct System A. With a system at
the design airflow (Q) and at a design system
resistance (P), an increase in airflow to 120% of Q
will result in an increase in system resistance P of
144% since system resistance varies with the square
of the airflow. Likewise, a decrease in airflow Q to
50% would result in a decrease in system resistance
P to 25% of the design system resistance.
In Figure 6.1, System Curve B is representative of a
system that has more component pressure loss than
System Curve A, and System Curve C has less
component pressure loss than System Curve A.
Pc/P = (Qc/Q)2
Notice that on a percentage basis, the same
relationships also hold for System Curves B and C.
These relationships are characteristic of typical fixed
systems.
A more through discussion of duct system pressure
losses can be found in AMCA Publication 200 Air
Systems.
200
160
140
SY
E
ST
M
C
120
100
EM
B
80
ST
S
SY
60
SY
PERCENT OF SYSTEM RESISTANCE
180
TE
M
A
SYSTEM
DESIGN
POINT
40
20
0
0
20
40
60
80
100
120
140
160
180
200
PERCENT OF SYSTEM AIRFLOW
Figure 6.1 - System Curves
Fans and Systems | 67
is now at Point 3 (the intersection of the fan curve and
the new System C), with the airflow at approximately
120% of Q.
6.4 Interaction of system curve and fan
performance curve
If the system characteristic curve, composed of the
resistance to system airflow and the appropriate SEF
have been accurately determined, then the fan will
deliver the designated airflow when installed in the
system.
6.5 Effect of changes in speed
Increases or decreases in fan rotational speed will
alter the airflow through a system. According to the
Fan Laws (see below), the % increase in airflow is
directly proportional to the fan rotational speed ratio,
and the fan static pressure is proportional to the
square of the fan rotational speed ratio. Thus, a 10%
increase in fan rotational speed will result in a new
fan curve with a 10% increase in Q, as illustrated in
Figure 6.3. Since the system components did not
change, System Curve A remains the same. With
airflow increasing by 10% over the original Q, the
system resistance increases along System Curve A
to Point 2, at the intersection with the new fan curve.
The point of intersection of the system curve and the
fan performance curve determines the actual airflow.
System Curve A in Figure 6.2 has been plotted with a
fan performance curve that intersects the system
design point.
The airflow through the system in a given installation
may be varied by changing the system resistance.
This is usually accomplished by using fan dampers,
duct dampers, mixing boxes, terminal units, etc.
The greater airflow moved by the fan against the
resulting higher system resistance to airflow is a
measure of the increased work done. In the same
system, the fan efficiency remains the same at all
points on the same system curve.
This is due to the fact that airflow, system resistance,
and required power are varied by the appropriate
ratio of the fan rotational speed.
A
200
ST
EM
180
160
SY
PERCENT OF SYSTEM RESISTANCE
Figure 6.2 shows the airflow may be reduced from
design Q by increasing the resistance to airflow, i.e.,
changing the system curve from System A to System
B. The new operating point is now at Point 2 (the
intersection of the fan curve and the new System B)
with the airflow at approximately 80% of Q. Similarly,
the airflow can be increased by decreasing the
resistance to airflow, i.e., changing the system curve
from System A to System C. The new operating point
140
SY
E
ST
M
C
FAN CURVE
2
120
1
100
80
SYSTEM
DESIGN
POINT
3
60
EM
B
T
YS
40
S
20
0
0
20
40
60
80
100
120
140
160
PERCENT OF SYSTEM AIRFLOW
Figure 6.2 - Interaction of System Curves and Fan Curve
68 | Fans and Systems
180
200
air density of 1.2 kg/m3 (0.075 lbm/ft3) is standard in
the fan industry throughout the world. Figure 6.4
illustrates the effect on the fan performance of a
density variation from the standard value.
6.5.1 Fan Laws - effect of change in speed - (fan
size and air density remaining constant)
For the same size fan, Dc = D and, therefore, (Dc/D)
= 1. When the air density does not vary, ρc = ρ and
the air density ratio (ρc/ρ) = 1. Kp is taken as equal to
unity in this and following examples.
6.6.1 Fan Laws - effect of change in density - (fan
size and speed remaining constant)
When the speed of the fan does not change, Nc = N
and, therefore, (Nc/N) = 1. The fan size is also fixed,
Dc = D and therefore (Dc/D) = 1.
Qc = Q × (Nc/N)
Ptc = Pt × (Nc/N)2
Psc = Ps × (Nc/N)2
Qc = Q
Pvc = Pv × (Nc/N)2
Ptc = Pt × (ρc/ρ)
Psc = Ps × (ρc/ρ)
Hc = H × (Nc/N)3
Pvc = Pv × (ρc/ρ)
6.6 Effect of density on system resistance
Hc = H × (ρc/ρ)
SY
CT
S (AT 1.1N)
160
PRESSURE
H (AT 1.1N)
133
140
S (AT N)
PRESSURE
120
2
H (AT N)
100
1
100
80
60
50
40
20
PERCENT OF POWER
DU
PERCENT OF SYSTEM RESISTANCE
ST
EM
A
The resistance of a duct system is dependent upon
the density of the air flowing through the system. An
110%
0
0
20
40
60
80
100
120
140
160
180
200
PERCENT OF SYSTEM AIRFLOW
Figure 6.3 - Effect of 10% increase in Fan Speed
Fans and Systems | 69
PERCENT OF SYSTEM
RESISTANCE AND FAN PRESSURE
FAN PRESSURE CURVE
@ DENSITY ρ
SYSTEM A
@ DENSITY ρ
FAN INLET
100
SYSTEM A
@ DENSITY ρ/2
FAN INLET
FAN PRESSURE CURVE
@ DENSITY ρ/2
80
60
40
20
0
PERCENT OF POWER
100
POWER @ DENSITY ρ
80
60
40
POWER @ DENSITY ρ/2
20
0
0
20
40
60
80
100
120
140
PERCENT OF SYSTEM AIRFLOW
Figure 6.4 - Density Effect
70 | Fans and Systems
160
180
200
6.7 Fan and system interaction
When system pressure losses have been accurately
estimated and desirable fan inlet and outlet
conditions have been provided, design airflow can be
expected, as illustrated in Figure 6.5. Note again that
the intersection of the actual system curve and the
fan curve determine the actual airflow. However,
when system pressure losses have not been
accurately estimated as in Figure 6.6, or when
undesirable fan inlet and outlet conditions exist as in
Figure 6.7, design performance may not be obtained.
6.8 Effects of errors in estimating system
resistance
6.8.1 Higher system resistance. In Figure 6.6,
System Curve B shows a situation where a system
has greater resistance to airflow than designed
(Curve A). This condition is generally a result of
inaccurate allowances of system resistance. All
pressure losses must be considered when
calculating system resistance or the actual system
will be more restrictive to airflow than intended. This
condition results in an actual airflow at Point 2, which
is at a higher pressure and lower airflow than was
expected.
If the actual duct system pressure loss is greater than
design, an increase in fan speed may be necessary
to achieve Point 5, the design airflow.
CAUTION: Before increasing fan rotational
speed, check with the fan manufacturer to
determine whether the fan rotational speed can
be safely increased. Also determine the expected
increase in power. Since the power required
increases as the cube of the fan rotational speed
ratio, it is very easy to exceed the capacity of the
existing motor and that of the available electrical
service.
6.8.2 Lower system resistance. Curve C in Figure
6.6 shows a system that has less resistance to airflow
than designed. This condition results in an actual
airflow at Point 3, which is at a lower pressure and
higher airflow than was expected.
CALCULATED SYSTEM CURVE
PEAK FAN PRESSURE
1
DESIGN RESISTANCE
FAN PRESSURE
CURVE
DESIGN AIRFLOW
Figure 6.5 - Fan/System Curve at Design Point
Fans and Systems | 71
the fan speed, adjusting the variable inlet vane (VIV),
if installed, or inlet dampers. The system resistance
could also be increased to Point 1 on Curve A, Figure
6.6. The change in fan operating point should be
evaluated carefully, since a change in fan power
consumption may occur.
6.9 Safety factors
It has been common practice among system
designers to add safety factors to the calculated
system resistance to account for the “unexpected”.
In some cases, safety factors may compensate for
resistance losses that were unaccounted for and the
actual system will deliver the design airflow, Point 1,
Figure 6.6. If the actual system resistance is lower
than the design system resistance, including the
safety factors, the fan will run at Point 3 and deliver
more airflow. This result may not be advantageous
because the fan may be operating at a less efficient
point on the fan’s performance curve and may require
more power than a properly designed system. Under
these conditions, it may be desirable to reduce the
fan performance to operate at Point 4 on Curve C,
Figure 6.6. This may be accomplished by reducing
The system designer should also evaluate the fan
performance tolerance and system resistance
tolerance to determine if the lower or upper limits of
the probable airflow in the system are acceptable.
The combination of these tolerances should be
evaluated to ensure that the “high-side” system
resistance curve does not fall into the unstable range
of performance. Operation in this area of the curve
should be avoided and precautions taken to ensure
operations outside of the unstable area, especially at
the highest expected system resistance.
CURVE B:
ACTUAL SYSTEM
ACTUAL SYSTEM RESISTANCE
MORE THAN DESIGN
CURVE A:
CALCULATED SYSTEM
5
CURVE C
ACTUAL SYSTEM
PEAK FAN
PRESSURE
2
1
DESIGN RESISTANCE
3
ACTUAL SYSTEM
LESS THAN
DESIGN
4
FAN PRESSURE
CURVE
DESIGN AIRFLOW
72 | Fans and Systems
Figure 6.6 - Fan/System Curve Not at Design Point
6.10 Deficient fan/system performance
The most common causes of deficient fan/system
performance are improper fan inlet duct design, fan
outlet duct design, and fan installation into the duct
system. Any one or a combination of these conditions
that alter the aerodynamic characteristics of the air
flowing through the fan such that the fan’s full airflow
potential, as tested in the laboratory and cataloged, is
not likely to be realized.
Other major causes of deficient performance are:
• The air performance characteristics of the
installed system are significantly different from
the system designer's intent (See Figure 6.6).
This may be due to a change in the system by
others or unexpected behavior of the system
during operation.
• The system design calculations did not include
adequate allowances for the effect of accessories
and appurtenances (See Section 10).
• The fan selection was made without allowing
for the effect of appurtenances on the fan's
performance (See Section 10).
• Dirty filters, dirty ducts, dirty coils, etc., will
increase the system resistance, and
consequently, reduce the airflow - often
significantly.
• The "performance" of the system has been
determined by field measurement techniques
that have a high degree of uncertainty.
Other "on-site" problems are listed in AMCA
Publication 202 Troubleshooting, which includes
detailed checklists and recommendations for the
correction of problems with the performance of air
systems.
6.11 Precautions
performance
to
prevent
deficient
• Use appropriate allowances in the design
calculations when space or other factors
dictate the use of less than optimum
arrangement of the fan outlet and inlet
connections (See Sections 8 and 9).
• Design the connections between the fan and
the system to provide, as nearly as possible,
uniform airflow conditions at the fan outlet and
inlet connections (See Sections 8 and 9).
• Include adequate allowance for the effect of all
accessories and appurtenances on the
performance of the system and the fan. If
possible, obtain from the fan manufacturer
data on the effect of installed appurtenances
on the fan's performance (See Section 10).
• Use field measurement techniques that can be
applied effectively on the particular system.
Be aware of the probable accuracy of
measurement and conditions that affect this.
Refer to AMCA Publication 203 Field
Performance Measurement of Fan Systems;
for more precise measurement see AMCA
Standard 803 Industrial Process/Power
Generation Fans: Site Performance Test
Standard. Also, refer to AABC National
Standards, Chapter 8, Volume Measurements,
Associated Air Balance Council, 5th Edition,
1989.
6.12 System Effect
Figure 6.7 illustrates deficient fan/system
performance resulting from one or more of the
undesirable airflow conditions listed in Section 6.10.
It is assumed that the system pressure losses, shown
in system curve A, have been accurately determined,
and a suitable fan selected for operation at Point 1.
However, no allowance has been made for the effect
of the system connections on the fan's performance.
To account for this System Effect it will be necessary
to add a System Effect Factor (SEF) to the calculated
system pressure losses to determine the actual
system curve. The SEF for any given configuration is
velocity dependent and will vary across a range of
airflow. This will be discussed in more detail in
Section 7. (See Figure 7.1).
In Figure 6.7 the point of intersection between the fan
performance curve and the actual system curve B is
Point 4. The actual airflow will be deficient by the
difference 1-4. To achieve design airflow, a SEF
equal to the pressure difference between Point 1 and
2 should have been added to the calculated system
pressure losses and the fan selected to operate at
Point 2. Note that because the System Effect is
velocity related, the difference represented between
Points 1 and 2 is greater than the difference between
Points 3 and 4.
The System Effect includes only the effect of the
system configuration on the fan's performance.
Fans and Systems | 73
7. System Effect Factor (SEF)
7.1 System Effect Curves
A System Effect Factor is a value that accounts for
the effect of conditions adversely influencing fan
performance when installed in the air system.
Figure 7.1 shows a series of 19 System Effect
Curves. By entering the chart at the appropriate air
velocity (on the abscissa), it is possible to read
across from any curve (to the ordinate) to find the
SEF for a particular configuration.
CURVE B
ACTUAL SYSTEM
WITH SYSTEM EFFECT
CURVE A
CALCULATED SYSTEM
WITH NO ALLOWANCE
FOR SYSTEM EFFECT
2
SYSTEM EFFECT LOSS
AT DESIGN AIRFLOW
4
DESIGN RESISTANCE
1
3
SYSTEM
EFFECT AT
ACTUAL AIRFLOW
FAN CATALOG
PRESSURE
CURVE
AIRFLOW
DEFICIENCY
DESIGN AIRFLOW
Figure 6.7 - Deficient Fan/System Performance - System Effect Ignored
74 | Fans and Systems
FG H I J K L
1000
M
N
O
P
900
Q
800
700
R
600
500
S
SYSTEM EFFECT FACTOR PRESSURE, Pa
400
300
T
U
200
V
100
90
W
80
70
60
X
50
40
30
20
2.5
3
4
5
6
7
8
9 10
20
30
AIR VELOCITY, (m/s)
(Air Density = 1.2 kg/m3)
Figure 7.1 - System Effect Curves (SI)
Fans and Systems | 75
FG H I J K L
M
N
O
5.0
P
4.0
Q
3.0
R
SYSTEM EFFECT FACTOR - PRESSURE, in. wg
2.5
2.0
S
1.5
T
1.0
0.9
U
0.8
0.7
0.6
V
0.5
0.4
W
0.3
0.25
X
0.2
0.15
0.1
5
6
7
8 9 10
15
20
25
30
AIR VELOCITY, ft/min × 100
(Air Density = 0.075 lbm/ft3)
Figure 7.1 - System Effect Curves (I-P)
76 | Fans and Systems
40
50
60
Table 7.1 - System Effect Coefficients
Curve in
Figure 7.1
Dynamic Pressure
Loss Coefficient C
F
G
H
I
J
K
L
M
N
O
P
Q
R
S
T
U
V
W
X
16.00
14.20
12.70
11.40
9.50
7.90
6.40
4.50
3.20
2.50
1.90
1.50
1.20
0.75
0.50
0.40
0.25
0.17
0.10
2
⎛ V ⎞
SEF = C ⎜
⎟ ρ
⎝ 1.414 ⎠
SI
2
⎛ V ⎞
SEF = C ⎜
⎟ ρ
⎝ 1097 ⎠
I-P
Fans and Systems | 77
The SEF is given in Pascals (in. wg) and must be
added to the total system pressure losses as shown
on Figure 7.2.
The velocity used when entering Figure 7.1 will be
either the inlet or the outlet velocity of the fan. This
will depend on whether the configuration in question
is related to the fan inlet or the fan outlet. Most
catalog ratings include outlet velocity figures but, for
centrifugal fans, it may be necessary to calculate the
inlet velocity (See Figure 9.14). The inlet velocity and
outlet velocity of an axial fan can be approximated by
using the fan impeller diameter to determine the
airflow area. The necessary dimensioned drawings
are usually included in the fan catalog.
In Sections 8 and 9, typical inlet and outlet
configurations are illustrated and the appropriate
System Effect Curve is listed for each configuration.
If more than one configuration is included in a
system, the SEF for each must be determined
separately and the total of these System Effects must
be added to the total pressure losses.
The System Effect Curves are plotted for standard air
at a density of 1.2 kg/m3 (0.075 lbm/ft3). Since the
System Effect is directly proportional to density,
values for other densities can be calculated as below:
⎛d ⎞
SEF2 = SEF1 ⎜ 2 ⎟
⎝ d1 ⎠
Where:
SEF2 = SEF at actual density
SEF1 = SEF at standard density
d2 = actual density
d1 = standard density
Alternatively, the SEF may be calculated by the
method shown in Table 7.1. Determine the
configuration being evaluated and use the
appropriate loss coefficient, Cp, and application
velocity, V. The SEF can then be calculated using the
equations shown in Table 7.1.
FAN POWER
ACTUAL SYSTEM RESISTANCE
ACTUAL POWER REQUIRED
ACTUAL SYSTEM W/ SEF
SEF
CALCULATED
SYSTEM W/NO
ALLOWANCE
FOR SEF
FAN PRESSURE
DESIGN AIRFLOW
Figure 7.2 - Effect of System on Fan Selection
78 | Fans and Systems
should examine catalog ratings carefully for
statements defining whether the published ratings
are based on tests made with A: free inlet, free outlet;
B: free inlet, ducted outlet; C: ducted inlet, free outlet
or D; ducted inlet, ducted outlet.
7.2 Power determination
When all the applicable System Effect Factors (SEF)
have been added to the calculated system pressure
losses the power shown in the catalog for the actual
point of operation, Figure 7.2 or Table 7.1 may be
used without further adjustment.
8.1 Outlet ducts
ANSI/AMCA 210 specifies an outlet duct that is no
greater than 105% or less than 95% of the fan outlet
area. It also requires that the slope of the transition
elements be no greater than 15° for converging
elements or greater than 7° for diverging elements.
As previously discussed, fans intended primarily for
use with duct systems are usually tested with an
outlet duct in place (See Figure 3.2). In most cases
it is not practical for the fan manufacturer to supply
this duct as part of the fan, but rated performance will
not be achieved unless a comparable duct is included
in the system design. The system design engineer
Figure 8.1 shows changes in velocity profiles at
various distances from centrifugal and axial flow fan
outlets. By definition, 100% "effective duct length" is
a minimum of two and one half (2½) equivalent duct
diameters. For velocities greater than 13 m/s (2500
fpm), add 1 duct diameter for each additional 5 m/s
(1000 fpm).
8. Outlet System Effect Factors
BLAST AREA
DISCHARGE DUCT
CUTOFF
OUTLET AREA
25%
50%
75%
CENTRIFUGAL FAN
100% EFFECTIVE DUCT LENGTH
AXIAL FAN
To calculate 100% duct length, assume a minimum of 2½ duct diameters for 12.7 m/s (2500 fpm) or less. Add 1
duct diameter for each additional 5.08 m/s (1000 fpm).
EXAMPLE: 25.4 m/s (5000 fpm) = 5 equivalent duct diameters. If the duct is rectangular with side dimensions a
and b, the equivalent duct diameter is equal to (4ab/π)0.5.
Figure 8.1 - Fan Outlet Velocity Profiles
Fans and Systems | 79
8.1.1 Axial flow fan - outlet ducts. Most exhaust
axial flow fans are tested and/or rated with two to
three equivalent duct diameters attached to the fan
outlet. Often, fans are installed without an outlet
duct, either because of available space or for
economic reasons. Tubeaxial fans installed with no
outlet ducts have System Effect Factors (SEF)
approaching zero.
Vaneaxial fans, however, do not perform as
cataloged when they are installed with less than 50%
"effective duct length." System Effect Curves for
tubeaxial and vaneaxial fans with less than optimum
outlet duct are shown in Figure 8.2.
To determine the applicable SEF, calculate the
average velocity in the outlet duct and enter the
System Effect Curve (Figure 7.1) at this velocity,
utilizing the appropriate System Effect Curve
selected from Figure 8.2, then read over horizontally
to the System Effect Factor, Pascals (in. wg) on the
ordinate.
8.1.2 Centrifugal flow fan - outlet ducts.
Centrifugal fans are sometimes installed with a less
than optimum outlet duct. If it is not possible to use a
full-length outlet duct, then a SEF must be added to
the system resistance losses. System Effect Curves
for centrifugal fans with less than optimum outlet duct
length are shown in Figure 8.3.
8.2 Outlet diffusers
Many air systems are space-constricted and must, of
necessity, use relatively small ducts having high
static pressure losses. If space is not severely
constricted, the use of larger ductwork and moving
air at a lower velocity may be beneficial. Larger
ductwork (within reason) reduces system pressure
requirements.
To effectively transition from a smaller duct size to a
larger duct size it is necessary to use a connection
piece between the duct sections that allows the
airstream to expand gradually. This piece is called a
diffuser, or evasé. These terms are used
interchangeably in the industry. A properly designed
evasé has a smooth and gradual transition between
the duct sizes so that airflow is relatively undisturbed.
An evasé operates on a very simple principle: air
flowing from the smaller area to the larger area loses
AXIAL FAN
100% EFFECTIVE DUCT LENGTH
To calculate 100% duct length, assume a minimum of 2½ duct diameters for 12.7 m/s (2500 fpm) or less. Add 1
duct diameter for each additional 5.08 m/s (1000 fpm).
EXAMPLE: 25.4 m/s (5000 fpm) = 5 equivalent duct diameters
No Duct
12%
Effective
Duct
25%
Effective
Duct
50 %
Effective
Duct
100%
Effective
Duct
Tubeaxial Fan
---
---
---
---
---
Vaneaxial Fan
U
V
W
---
---
Determine SEF by using Figure 7.1
Figure 8.2 - System Effect Curves for Outlet Ducts - Axial Fans
80 | Fans and Systems
velocity as it approaches the larger area, and a
portion of the change (reduction) in velocity pressure
is converted into static pressure. This process is
called “static regain”, and is simply defined as the
conversion of velocity pressure to static pressure.
The efficiency of conversion (or loss of total pressure)
will depend upon the angle of expansion, the length
of the evasé section, and the blast area/outlet area
ratio of the fan.
The fan manufacturer will, in most cases, be able to
provide design information for an efficient diffuser.
See AMCA Publication 200 Air Systems, for an
example showing the effect of a diffuser on a duct
exit.
8.3 Outlet duct elbows
Values for pressure losses through elbows, which are
published in handbooks and textbooks, are based
upon a uniform velocity profile at entry into the elbow.
Any non-uniformity in the velocity profile ahead of the
elbow will result in a pressure loss greater than the
industry-accepted value.
BLAST AREA
DISCHARGE DUCT
OUTLET AREA
CUTOFF
100% EFFECTIVE DUCT LENGTH
CENTRIFUGAL FAN
To calculate 100% duct length, assume a minimum of 2½ duct diameters for 2500 fpm or less. Add 1 duct diameter
for each additional 1000 fpm.
EXAMPLE: 5000 fpm = 5 equivalent duct diameters. If the duct is rectangular with side dimensions a and b, the
equivalent duct diameter is equal to (4ab/π)0.5.
Pressure
Recovery
No Duct
12%
Effective Duct
25%
Effective Duct
50%
Effective Duct
100%
Effective Duct
0%
50%
80%
90%
100%
W
W
W-X
—
—
—
—
—
—
—
—
—
—
—
Blast Area
Outlet Area
0.4
0.5
0.6
0.7
0.8
0.9
1.0
System Effect Curve
P
P
R-S
S
T-U
V-W
—
R-S
R-S
S-T
U
V-W
W-X
—
U
U
U-V
W-X
X
—
—
Determine SEF by using Figure 7.1
Figure 8.3 - System Effect Curves for Outlet Ducts - Centrifugal Fans
Fans and Systems | 81
Since the velocity profile at the outlet of a fan is not
uniform, an elbow located at or near the fan outlet will
develop a pressure loss greater than the industryaccepted value.
8.3.1 Axial fans - outlet duct elbows. Tubeaxial
fans with two-piece and four-piece mitered elbows at
varying distances from the fan outlet have a
negligible SEF (see Figure 8.4).
The amount of this loss will depend upon the location
and orientation of the elbow relative to the fan outlet.
In some cases, the effect of the elbow will be to
further distort the outlet velocity profile of the fan.
This will increase the losses and may result in such
uneven airflow in the duct that branch- takeoffs near
the elbow will not deliver their design airflow. (See
Section 8.6)
Vaneaxial fans with two and four-piece mitered
elbows at varying distances from the fan outlet
resulted in System Effect Curves as shown in Figure
8.4.
Wherever possible, a length of straight duct should
be installed at the fan outlet to permit the diffusion
and development of a uniform airflow profile before
an elbow is inserted in the duct. If an elbow must be
located near the fan outlet then it should be a radius
elbow having a minimum radius-to-duct-diameter
ratio of 1.5.
8.3.2 Centrifugal fans - outlet duct elbows. The
outlet velocity of centrifugal fans is generally higher
toward one or adjacent sides of the rectangular duct.
If an elbow must be located near the fan outlet it
should have a minimum radius-to-duct-diameter ratio
of 1.5, and it should be arranged to give the most
uniform airflow possible. Figure 8.5 gives System
Effect Curves that can be used to estimate the effect
of an elbow at the fan outlet. It also shows the
reduction in losses resulting from the use of a straight
outlet duct.
TUBEAXIAL FAN SHOWN
% EFFECTIVE
DUCT LENGTH
% EFFECTIVE
DUCT LENGTH
VANEAXIAL FAN SHOWN
90° Elbow
No Duct
12%
Effective
Duct
25%
Effective
Duct
50 %
Effective
Duct
100%
Effective
Duct
Tubeaxial Fan
2 & 4 Pc
---
---
---
---
---
Vaneaxial Fan
2 Pc
U
U-V
V
W
---
Vaneaxial Fan
4 Pc
W
---
---
---
---
Determine SEF by using Figures 7.1 and 8.1
Figure 8.4 - System Effect Curves for Outlet Duct Elbows - Axial Fans
82 | Fans and Systems
POSITION C
POSITION D
POSITION B
E
TIV TH
C
G
FE
EF LEN
% CT
DU
INL
ET
POSITION A
SWSI CENTRIFUGAL FAN SHOWN
Note: Fan Inlet and elbow positions must be oriented as shown for the proper application of the table on the facing
page.
Figure 8.5 - Outlet Elbows on SWSI Centrifugal Fans
Fans and Systems | 83
Outlet
Elbow
Position
No Outlet
Duct
12%
Effective
Duct
25%
Effective
Duct
50%
Effective
Duct
0.4
A
B
C
D
N
M-N
L-M
L-M
O
N
M
M
P-Q
O-P
N
N
S
R-S
Q
Q
0.5
A
B
C
D
O-P
N-O
M-N
M-N
P-Q
O-P
N
N
R
Q
O-P
O-P
T
S-T
R-S
R-S
0.6
A
B
C
D
Q
P
N-O
N-O
Q-R
Q
O
O
S
R
Q
Q
U
T
S
S
0.7
A
B
C
D
R-S
Q-R
P
P
S
R-S
Q
Q
T
S-T
R-S
R-S
V
U-V
T
T
0.8
A
B
C
D
S
R-S
Q-R
Q-R
S-T
S
R
R
T-U
T
S
S
W
V
U-V
U-V
0.9
A
B
C
D
T
S
R
R
T-U
S-T
S
S
U-V
T-U
S-T
S-T
W
W
V
V
1.0
A
B
C
D
T
S-T
R-S
R-S
T-U
T
S
S
U-V
U
T
T
W
W
V
V
SYSTEM EFFECT CURVES FOR SWSI FANS
DETERMINE SEF BY USING FIGURES 7.1 AND 8.1
For DWDI fans determine SEF using the curve for SWSI
fans. Then, apply the appropriate multiplier from the
tabulation below
MULTIPLIERS FOR DWDI FANS
ELBOW POSITION A = ∆P × 1.00
ELBOW POSITION B = ∆P × 1.25
ELBOW POSITION C = ∆P × 1.00
ELBOW POSITION D = ∆P × 0.85
Figure 8.5 - Outlet Elbows on SWSI Centrifugal Fans
84 | Fans and Systems
100%
Effective
Duct
NO System Effect Factor
Blast Area
Outlet Area
a large plenum or to free space a parallel blade
damper may be satisfactory.
8.4 Turning vanes
Turning vanes will usually reduce the pressure loss
through an elbow, however, where a non-uniform
approach velocity profile exists, such as at a fan
outlet, the vanes may serve to continue the nonuniform
profile
beyond
the
elbow.
This may result in increased losses in other system
components downstream of the elbow.
8.5 Volume control dampers
Volume control dampers are manufactured with
either "opposed" blades or "parallel" blades. When
partially closed, the parallel bladed damper diverts
the airstream to the side of the duct.
This results in a non-uniform velocity profile beyond
the damper and airflow to branch ducts close to the
downstream side may be seriously affected.
The use of an opposed blade damper is
recommended when air volume control is required at
the fan outlet and there are other system
components, such as coils or branch takeoffs
downstream of the fan. When the fan discharges into
PARALLEL-BLADE DAMPER
ILLUSTRATING DIVERTED AIRFLOW
For a centrifugal fan, best air performance will usually
be achieved by installing an opposed blade damper
with its blades perpendicular to the fan shaft;
however, other considerations, such as the need for
thrust bearings, may require installation of the
damper with its blades parallel to the fan shaft.
When a damper is required, it is often furnished as
accessory equipment by the fan manufacturer (see
Figure 8.6). In many systems, a volume control
damper will be located in the ductwork at or near the
fan outlet.
Published pressure drops for wide-open control
dampers are based on uniform approach velocity
profiles. When a damper is installed close to the
outlet of a fan the approach velocity profile is nonuniform and much higher pressure losses through the
damper can result. Figure 8.7 lists multipliers that
should be applied to the damper manufacturer's
catalog pressure drop when the damper is installed at
the outlet of a centrifugal fan. These multipliers
should be applied to all types of fan outlet dampers.
OPPOSED-BLADE DAMPER
ILLUSTRATING NON-DIVERTED
AIRFLOW
Figure 8.6 - Parallel Blade vs. Opposed Blade Damper
Fans and Systems | 85
VOLUME CONTROL DAMPER
BLAST AREA
OUTLET AREA
PRESSURE DROP
MULTIPLIER
0.4
7.5
0.5
4.8
0.6
3.3
0.7
2.4
0.8
1.9
0.9
1.5
1.0
1.2
Figure 8.7 - Pressure Drop Multipliers for Volume Control Dampers on a Fan Discharge
86 | Fans and Systems
8.6 Duct branches
Standard procedures for the design of duct systems
are based on the assumption of uniform airflow
profiles in the system.
In Figure 8.8 branch takeoffs or splits are located
close to the fan outlet. Non-uniform airflow conditions
will exist and pressure loss and airflow may vary
widely from the design intent. Wherever possible a
length of straight duct should be installed between
the fan outlet and any split or branch takeoff.
Note: Avoid location of split or duct branch close to fan discharge. Provide a straight section of duct to allow for air
diffusion.
Figure 8.8 - Branches Located Too Close to Fan
Fans and Systems | 87
loss of energy, or even a flat flange (e) on the end of
the duct or fan will reduce the loss to about one half
of the loss through an un-flanged entry.
9. Inlet System Effect Factors
Fan performance can be greatly affected by nonuniform or swirling inlet flow. Fan rating and catalog
performance is typically obtained with unobstructed
inlet flow. Any disruption to the inlet airflow will reduce
a fan’s performance. Restricted fan inlets located
close to walls, obstructions or restrictions caused by
a plenum or cabinet will also decrease the
performance of a fan. The fan performance loss due
to inlet airflow disruption must be considered as a
System Effect.
ANSI/AMCA 210 limits an inlet duct to a crosssectional area no greater than 112.5% or less than
92.5% of the fan inlet area. The slope of transition
elements is limited to 15° converging and 7° diverging.
9.2 Inlet duct elbows
Non-uniform airflow into a fan inlet is a common
cause of deficient fan performance. An elbow located
at, or in close proximity to the fan inlet will not allow
the air to enter the impeller uniformly. The result is
less than cataloged air performance.
9.1 Inlet ducts
Fans intended primarily for use as "exhausters" may
be tested with an inlet duct in place, or with a special
bell-mouthed inlet to simulate the effect of a duct.
Figure 9.1 illustrates variations in inlet airflow that will
occur. The ducted inlet condition is shown as (a), and
the effect of the bell-mouth inlet as (b).
A word of caution is required with the use of inlet
elbows in close proximity to fan inlets. Other than the
incurred System Effect Factor, instability in fan
operation may occur as evidenced by an increase in
pressure fluctuations and sound power level. Fan
instability, for any reason, may result in serious
structural damage to the fan. Axial fan instabilities
were experienced in some configurations tested with
inlet elbows in close proximity to the fan inlet.
Pressure fluctuations approached ten (10) times the
magnitude of fluctuations of the same fan with good
inlet and outlet conditions. It is strongly advised
that inlet elbows be installed a minimum of three
(3) diameters away from any axial or centrifugal
fan inlet.
Fans that do not have smooth entries (c), and are
installed without ducts, exhibit airflow characteristics
similar to a sharp edged orifice that develops a vena
contracta. A reduction in airflow area is caused by the
vena contracta and the following rapid expansion
causes a loss that should be considered as a System
Effect.
If it is not practical to include such a smooth entry, a
converging taper (d) will substantially diminish the
c.
a.
b.
IDEAL SMOOTH ENTRY TO
DUCT ON A DUCT SYSTEM
BELL MOUTH INLET PRODUCES
FULL FLOW INTO FAN
d.
CONVERGING TAPERED ENTRY
INTO FAN OR DUCT SYSTEM
VENA CONTRACTA AT INLET
REDUCES EFFECTIVE FAN INLET AREA
e.
FLANGED ENTRY INTO
FAN OR DUCT SYTEM
Figure 9.1 Typical Inlet Connections for Centrifugal and Axial Fans
88 | Fans and Systems
9.2.1 Axial fans - inlet duct elbows. The System
Effect Curves shown in Figure 9.2 for tubeaxial and
vaneaxial fans are the result of tests run with two and
four piece mitered inlet elbows at or in close proximity
to the fan inlets. Other variables tested included hubto-tip (H/T) ratio and blade solidity. The number of
blades did not have a significant affect on the inlet
elbow SEF.
are listed on Figure 9.4, and the System Effect
Curves for various square duct elbows of given
radius/diameter ratios are listed on Figure 9.5. The
SEF for a particular elbow is found in Figure 7.1 at
the intersection of the average fan inlet velocity and
the tabulated System Effect Curve.
This pressure loss should be added to the friction and
dynamic losses already determined for the particular
elbow. Note that when duct turning vanes and/or a
suitable length of duct is used (three to eight
diameters long, depending on velocities) between the
fan inlet and the elbow, the SEF is not as great.
These improvements help maintain uniform airflow
9.2.2 Centrifugal fans - inlet duct elbows. Nonuniform airflow into a fan inlet, Figure 9.3A, is a
common cause of deficient fan performance.
The System Effect Curves for mitered 90° round
section elbows of given radius/diameter (R/D) ratios
TUBEAXIAL FAN SHOWN
DUCT LENGTH
DUCT LENGTH
VANEAXIAL FAN SHOWN
H/T
90° Elbow
No Duct [1][2]
0.5D [1][2]
1.0D [1][2]
3.0D
Tubeaxial Fan
.25
2 piece
U
V
W
---
Tubeaxial Fan
.25
4 piece
X
---
---
---
Tubeaxial Fan
.35
2 piece
V
W
X
Vaneaxial Fan
.61
2 piece
Q-R
Q-R
S-T
T-U
Vaneaxial Fan
.61
4 piece
W
W-X
---
---
Notes:
[1] Instability in fan operation may occur as evidenced by an increase in pressure fluctuations and sound level.
Fan instability, for any reason, may result in serious structural damage to the fan.
[2] The data presented in Figure 9.2 is representative of commercial type tubeaxial and vaneaxial fans, i.e. 60%
to 70% fan static efficiency.
Figure 9.2 - System Effect Curves for Inlet Duct Elbows - Axial Fans
Fans and Systems | 89
into the fan inlet and thereby approach the airflow
conditions of the laboratory test setup.
Occasionally, where space is limited, the inlet duct
will be mounted directly to the fan inlet as shown on
Figure 9.3B. The many possible variations in the
width and depth of a duct influence the reduction in
performance to varying degrees and makes it
impossible to establish reliable SEF. Note: Capacity
losses as high as 45% have been observed in
poorly designed inlets such as in Figure 9.3B.
This inlet condition should be AVOIDED.
Existing installations can be improved with guide
vanes or the conversion to square or mitered elbows
with guide vanes, but a better alternative would be a
specially designed inlet box similar to that shown in
Figure 9.6.
9.2.3 Inlet boxes. Inlet boxes are added to
centrifugal and axial fans instead of elbows in order
to provide more predictable inlet conditions and to
maintain stable fan performance. They may also be
used to protect fan bearings from high temperature,
or corrosive / erosive gases. The fan manufacturer
should include the effect of any inlet box on the fan
performance, and when evaluating a proposal it
should be established that an appropriate loss has
been incorporated in the fan rating. Should this
information not be available from the manufacturer,
refer to Section 10.4 for an approximate System Effect.
A counter-rotating vortex at the inlet may result in a
slight increase in the pressure-volume curve but the
power will increase substantially.
There are occasions, with counter-rotating swirl,
when the loss of performance is accompanied by a
surging airflow. In these cases, the surging may be
more objectionable than the performance change.
Inlet spin may arise from a great variety of approach
conditions and sometimes the cause is not obvious.
D
LENGTH
OF DUCT
R
Figure 9.3A - Non-Uniform Airflow Into a Fan
Inlet Induced by a 90°, 3-Piece Section Elbow-No Turning Vanes
9.3 Inlet vortex (spin or swirl)
Another major cause of reduced performance is an
inlet duct design or fan installation that produces a
vortex or spin in the airstream entering a fan inlet.
An example of this condition is illustrated in Figure
9.7.
An ideal inlet condition allows the air to enter
uniformly without spin in either direction.
A spin in the same direction as the impeller rotation
(pre-rotation) reduces the pressure- volume curve by
an amount dependent upon the intensity of the
vortex. The effect is similar to the change in the
pressure-volume curve achieved by variable inlet
vanes installed in a fan inlet; the vanes induce a
controlled spin in the direction of impeller rotation,
reducing the airflow, pressure and power (see
Section 10.6).
90 | Fans and Systems
Figure 9.3B - Non-Uniform Airflow Induced Into
Fan Inlet by a Rectangular Inlet Duct
SYSTEM EFFECT CURVES
LENGTH
OF DUCT
D
R/D
NO
DUCT
—
N
+
R
2D
5D
DUCT DUCT
P
R-S
Figure 9.4A - Two Piece Mitered 90° Round Section Elbow - Not Vaned
SYSTEM EFFECT CURVES
LENGTH
OF DUCT
D
R/D
NO
DUCT
2D
5D
DUCT DUCT
0.5
O
Q
S
0.75
Q
R-S
T-U
1.0
R
S-T
U-V
2.0
R-S
T
U-V
3.0
S
T-U
V
+
R
Figure 9.4B - Three Piece Mitered 90° Round Section Elbow - Not Vaned
SYSTEM EFFECT CURVES
LENGTH
OF DUCT
R/D
NO
DUCT
2D
5D
DUCT DUCT
0.5
P-Q
R-S
T
0.75
Q-R
S
U
1.0
R
S-T
U-V
2.0
R-S
T
U-V
3.0
S-T
U
V-W
D
+
R
Figure 9.4C - Four or More Piece Mitered 90° Round Section Elbow - Not Vaned
DETERMINE SEF BY USING FIGURE 7.1
Figure 9.4 - System Effect Curves for Various Mitered Elbows without Turing Vanes
Fans and Systems | 91
SYSTEM EFFECT CURVES
H
R/D
NO
DUCT
2D
5D
DUCT DUCT
0.5
O
Q
S
0.75
P
R
S-T
1.0
R
S-T
U-V
1.0
S
T-U
V
H
LENGTH
OF DUCT
+
R
Figure 9.5A - Square Elbow with Inlet Transition - No Turning Vanes
H
SYSTEM EFFECT CURVES
R/D
NO
DUCT
0.5
S
T-U
V
1.0
T
U-V
W
2.0
V
V-W
W-X
H
LENGTH
OF DUCT
+
2D
5D
DUCT DUCT
R
Figure 9.5B - Square Elbow with Inlet Transition - 3 Long Turning Vanes
SYSTEM EFFECT CURVES
H
R/D
NO
DUCT
0.5
S
T-U
V
1.0
T
U-V
W
2.0
V
V-W
W-X
H
LENGTH
OF DUCT
R
+
2D
5D
DUCT DUCT
Figure 9.5C - Square Elbow with Inlet Transition - Short Turning Vanes
D = Diameter of the inlet collar
The inside area of the square duct (H x H) should be equal to the inside area of the fan inlet collar.
* The maximum permissible angle of any converging element of the transition is 15°, and for a diverging element, 7°.
DETERMINE SEF BY USING FIGURE 7.1
Figure 9.5 - System Effect Curves for Various Square Duct Elbows
92 | Fans and Systems
Figure 9.6 - Improved Flow Conditions with a Special Designed Inlet Box
IMPELLER
ROTATION
COUNTER-ROTATING SWIRL
Figure 9.7 - Example of a Forced Inlet Vortex
IMPELLER
ROTATION
PRE-ROTATING SWIRL
IMPELLER
ROTATION
COUNTER-ROTATING SWIRL
Figure 9.8 - Inlet Duct Connections Causing Inlet Spin
Fans and Systems | 93
airflow entering a duct elbow with turning vanes will
leave the duct elbow with non-uniform airflow.
9.4 Inlet turning vanes
Where space limitations prevent the use of optimum
fan inlet conditions, more uniform airflow can be
achieved by the use of turning vanes in the inlet
elbow (see Figure 9.9). Numerous variations of
turning vanes are available, from a single curved
sheet metal vane to multi-bladed "airfoil" vanes.
The pressure drop (loss) through these devices must
be added to the system pressure losses.
The amount of loss for each device is published by
the manufacturer, but it should be realized that the
cataloged pressure loss will be based upon uniform
airflow at the entry to the elbow. If the airflow
approaching the elbow is significantly non-uniform
because of a disturbance farther upstream in the
system, the pressure loss through the elbow will be
higher than the published figure. A non-uniform
9.5 Airflow straighteners
Figure 9.10 shows two airflow straighteners used in
testing setups to reduce fan swirl before measuring
stations. Figure 9.10A is the egg-crate straightener
used in ANSI/AMCA 210; larger cell sizes made
proportionately longer could be used.
Figure 9.10B shows the star straightener used in the
ISO standard. A single splitter sheet may be used to
eliminate swirl in some cases. Straighteners are
intended to reduce swirl before or after a fan or a
process station. Do not install straighteners where
the air profile is known to be non-uniform, the
device will carry the non-uniformity further
downstream.
TURNING
VANES
TURNING
VANES
IMPELLER
ROTATION
CORRECTED PREROTATING SWIRL
TURNING
VANES
CORRECTED COUNTERROTATING SWIRL
Figure 9.9 - Corrections for Inlet Spin
94 | Fans and Systems
IMPELLER
ROTATION
0.45D
D
0.075D
DUCT
0.075D
Figure 9.10A - ANSI/AMCA Standard 210 Egg-Crate Straightener
DUCT
DUCT
D
2D
Figure 9.10B - ISO 5801 Star Straightener
Figure 9.10 - Test Standard Airflow Straighteners
Fans and Systems | 95
one-half impeller diameter between an enclosure wall
and the fan inlet. Adjacent inlets of multiple double
width centrifugal fans located in a common enclosure
should be at least one impeller diameter apart if
optimum performance is to be expected. Figure 9.11
illustrates fans with restricted inlets and their
applicable System Effect Curves.
9.6 Enclosures (plenum and cabinet effects)
Fans within plenums and cabinets or next to walls
should be located so that air may flow unobstructed
into the inlets. Fan performance is reduced if the
space between the fan inlet and the enclosure is too
restrictive. It is common practice to allow at least
2L
L
EQUAL
INLET
DIA.
L
EQUAL
DIAMETER
OF INLET
Figure 9.11A - Fans and Plenum
L
L
Figure 9.11B - Axial Fan Near Wall
L
DWDI
L
SWSI
Figure 9.11C - Centrifugal Fan Near Wall(s)
Figure 9.11D - DWDI Fan Near Wall on One Side
L - DISTANCE
INLET TO WALL
For Figures 9.11A, B & C
SYSTEM EFFECT CURVES
For Figures 9.11D
SYSTEM EFFECT CURVES
0.75 x DIA. OF INLET
0.50 x DIA. OF INLET
0.40 x DIA. OF INLET
0.30 x DIA. OF INLET
V-W
U
T
S
X
V-W
V-W
U
Determine SEF by calculating inlet velocity and using Figure 7.1
Figure 9.11 - System Effect Curves for Fans Located in Plenums and Cabinet
Enclosures and for Various Wall-to-inlet Dimensions
96 | Fans and Systems
L
The manner in which the air stream enters an
enclosure in relation to the fan inlets also affects fan
performance. Plenum or enclosure inlets or walls that
are not symmetrical with the fan inlets will cause
uneven airflow and/or inlet spin. Figure 9.12A
illustrates this condition that must be avoided to
achieve maximum performance from a fan. If this is
not possible, inlet conditions can usually be improved
with a splitter sheet to break up the inlet vortex as
illustrated in Figure 9.12B.
common inlet obstructions. Some accessories such
as fan bearings, bearing pedestals, inlet vanes, inlet
dampers, drive guards and motors may also cause
inlet obstruction and are discussed in more detail in
Section 10.
Obstruction at the fan inlet may be defined in terms
of the unobstructed percentage of the inlet area.
Because of the shape of the inlet cones of many fans
it is sometimes difficult to establish the area of the fan
inlet. Figure 9.14 illustrates the convention adopted
for this purpose. Where an inlet collar is provided, the
inlet area is calculated from the inside diameter of
this collar. Where no collar is provided, the inlet plane
is defined by the points of tangency of the fan
housing side with the inlet cone radius.
For proper performance of axial fans in parallel
installations minimum space of one impeller diameter
should be allowed between fans, as shown in Figure
9.13. Placing fans closer together can result in erratic
or uneven airflow into the fans.
The unobstructed percentage of the inlet area is
calculated by projecting the profile of the obstruction
on the profile of the inlet. The adjusted inlet velocity
obtained is then used to enter the System Effect Curve
chart and the SEF determined from the curve listed
for that unobstructed percentage of the fan inlet area.
9.7 Obstructed inlets
A reduction in fan performance can be expected
when an obstruction to airflow is located in the fan
inlet. Building structural members, columns, butterfly
valves, blast gates and pipes are examples of more
SPLITTER SHEET
Figure 9.12A - Enclosure Inlet Not
Symmetrical with Fan Inlet. PreRotational Vortex Induced
Figure 9.12B - Flow Condition of Figure 9.12A
Improved with a Splitter Sheet. Substantial
Improvement Would Be To Relocate
Enclosure Inlet as Shown in Figure 9.11A
Figure 9.12 - Fan in Plenum with Non-Symmetrical Inlet
1 DIA.
MIN
Figure 9.13 - Parallel Installation of Axial Flow Fans
Fans and Systems | 97
ER
ET
AM
DE
SI
IN
DI
T
AR
LL
CO
E
NL
I
INLET PLANE
FREE INLET AREA PLANE - FAN WITH INLET COLLAR
POINT OF TANGENT
WITH FAN HOUSING SIDE
AND INLET CONE RADIUS
R
TE
E
NT
AM
GE
N
DI
TA
OF
INLET PLANE
FREE INLET AREA PLANE - FAN WITHOUT INLET COLLAR
Table for Figure 9.14
System Effect Curve (Figure 7.1)
Distance from obstruction to inlet plane
Percentage of
unobstructed inlet area
0.75 Inlet
diameter
0.5 Inlet
diameter
0.33 Inlet
diameter
0.25 Inlet
diameter
At Inlet plane
100
-
-
-
-
-
95
-
-
X
W
V
90
-
X
V-W
U-V
T-U
85
X
W-X
V-W
U-V
S-T
75
W-X
V
U
S-T
R-S
50
V-W
U
S-T
R-S
Q
25
U-V
T
S-T
Q-R
P
Figure 9.14 - System Effect Curves for Inlet Obstructions
(Table based on Fans and Fan Systems, Thompson & Trickler, Chem Eng MAR83, p. 60)
98 | Fans and Systems
10. Effects of Factory Supplied Accessories
Unless the manufacturer's catalog clearly states to
the contrary, it should be assumed that published fan
performance data does not include the effects of any
accessories supplied with the fan.
If possible, the necessary information should be
obtained directly from the manufacturer. The data
presented in this section are offered only as a guide
in the absence of specific data from the fan
manufacturer. See Figure 10.1 for terminology.
Cone Type
Variable
Inlet Vanes
Figure 10.1 - Common Terminology for Centrifugal Fan Appurtenances
Fans and Systems | 99
10.1 Bearing and supports in fan inlet
10.3 Belt tube in axial fan inlet or outlet
Arrangement 3 and 7 fans (see Figure 3.5) require
that the fan shaft be supported by a bearing and
bearing support in the fan inlet or just adjacent to it.
With a belt driven axial flow fan it is usually necessary
that the fan motor be mounted outside the fan
housing (see Figure 3.7 Arrangement 9, and Annex B
Figure B.7).
These components may have an effect on the flow of
air into the fan inlet and consequently on the fan
performance, depending upon the size of the
bearings and supports in relation to the fan inlet
opening. The location of the bearing and support, that
is, whether it is located in the actual inlet or "spaced
out" from the inlet, will also have an effect.
In cases where manufacturer's performance ratings
do not include the effect of the bearings and
supports, it will be necessary to compensate for this
inlet restriction. Use the fan manufacturer's
allowance for bearings in the fan inlet if possible.
If no better data are available, use the procedures
described in Section 9.7 as an approximation.
To protect the belts from the airstream, and also to
prevent any air leakage through the fan housing,
manufacturers in many cases provide a belt tube.
Most manufacturers include the effects of an axial fan
belt tube in their rating tables. In cases where the
effect is not included, the appropriate SEF is
approximated by calculating the percentage of
unobstructed area of air passage way and using
Figure 9.14.
10.4 Inlet box
When an inlet box configuration is supplied by the fan
manufacturer, the fan performance should include
the effect of the inlet box.
10.2 Drive guards obstructing fan inlet
All fans have moving parts that require guarding for
safety in the same way as other moving machinery.
Fans located less than 2.1 m (7 ft) above the floor
require special consideration as specified in the
United States’ Occupational Safety and Health Act.
National, federal, state and local rules, regulations,
and codes should be carefully considered and
followed.
Arrangement 3 and 7 fans may require a belt drive
guard in the area of the fan inlet. Depending on the
design, the guard may be located in the plane of the
inlet, along the casing side sheet, or it may be
"spaced out" due to "spaced out" bearing pedestals.
In any case, depending on the location of the guard,
and on the inlet velocity, the fan performance may be
significantly affected by this obstruction. It is
desirable that a drive guard located in this position be
furnished with as much opening as possible to allow
maximum flow of air to the fan inlet.
If available, use the fan manufacturer's allowance for
drive guards obstructing the fan inlet. SEF for drive
guard obstructions situated at the inlet of a fan may
be approximated using Figure 9.14.
Where possible, open construction on guards is
recommended to allow free air passage to the fan
inlet. Guards and sheaves should be designed to
obstruct, as little of the fan inlet as possible and in no
case should the obstruction be more than 1/3 of the
fan inlet area.
100 | Fans and Systems
The System Effect of fan inlet boxes can vary widely
depending upon the design. This data should be
available from the fan manufacturer. In the absence
of fan manufacturer's data, a well-designed inlet box
should approximate System Effect Curves "S" or "T"
of Figure 7.1.
10.5 Inlet box dampers
Inlet box dampers may be used to control the airflow
through the system. Either parallel or opposed blades
may be used (see Figure 10.1).
The parallel blade type is installed with the blades
parallel to the fan shaft so that, in a partially closed
position, a forced inlet vortex will be generated. The
effect on the fan characteristics will be similar to that
of a variable inlet vane control.
The opposed blade type is used to control airflow by
the addition of pressure loss created by the damper
in a partially closed position.
If possible, complete data should be obtained from
the fan manufacturer giving the System Effect of the
inlet box and damper pressure drop over the range of
application. If data are not available, System Effect
Curves "S" or "T" from Figure 7.1 should be applied
for the inlet box and pressure loss from the damper
manufacturer for the damper in making the fan
selection.
When variable inlet vanes are supplied by the fan
manufacturer, the performance should include the
effects of the variable inlet vane unit.
10.6 Variable inlet vane (VIV)
Variable inlet vanes are mounted on the fan inlet to
maintain fan efficiency at reduced airflow. They are
arranged to generate an inlet vortex (pre-rotation)
that rotates in the same direction as the fan impeller.
Variable inlet vanes may be of two different basic
types: 1) cone type integral with the fan inlet, 2)
cylindrical type add-on (Figures 10.1 and 10.2).
VANE TYPE
The System Effect of a wide-open VIV (see Figure
10.2) must be accounted for in the original fan
selection. If data are not available from the fan
manufacturer the following System Effect Curves
should be applied in making the fan selection.
SYSTEM EFFECT CURVE
(100% Open)
a) Cone type, integral
b) Cylindrical type
“Q” or “R”
“S”
Determine SEF by calculating inlet velocity and using
Figure 7.1
FAN PERFORMANCE
W/OUT VARIABLE INLET VANES
120
CONE TYPE
VARIABLE INLET
VANES
CYLINDRICAL TYPE
VARIABLE INLET
VANES
PERCENT OF SHUT-OFF PRESSURE
100
VARIABLE
INLET VANES
100% OPEN
75% OPEN
80
75% OPEN
60
40
75% OPEN
20
0
20
40
60
80
100
120
PERCENT OF WIDE OPEN VOLUME
Figure 10.2 - Typical Variable Inlet Vanes for a Backward Inclined Fan
Fans and Systems | 101
Annex A. SI / I-P Conversion Table (Informative)
Taken from AMCA 99-0100
Quantity
I-P to SI
SI to I-P
Length
(ft) 0.3048 = m
(m) 3.2808 = ft
Mass (weight)
(lbs) 0.4536 = kg
(kg) 2.2046 = lbs.
Time
The unit of time is the second in both systems
Velocity
(ft-s) 0.3048 = ms
(ft/min) 0.00508 = ms
(ms) 3.2808 = ft-s
(ms) 196.85 = ft/min
Acceleration
(in./s2) 0.0254 = m/s2
(m/s2) 39.370 = in./s2
Area
(ft2) 0.09290 = m2
(m2) 10.764 = ft2
Volume Flow Rate
(cfm) 0.000471948 = m3/s
(m3/s) 2118.88 = cfm
Density
(lb/ft3) 16.01846 = kg/m3
(kg/m3) 0.06243 = lb/ft3
Pressure
(in. wg) 248.36 = Pa
(in. wg) 0.24836 = kPa
(in. Hg) 3.3864 = kPa
(Pa) 0.004026 = in. wg
(kPa) 4.0264 = in. wg
(kPa) 0.2953 = in. Hg
Viscosity:
Absolute
Kinematic
(lbm/ft-s) 1.4882 = Pa s
(ft2/s) 0.0929 = m2/s
(Pa s) 0.6719 = (lbm/ft-s)
(m2/s) 10.7639 = ft2/s
Gas Constant
(ft lb/lbm-°R) 5.3803 = J-kg/K
(j-kg/K) 0.1858 = (ft lb/lbm-°R)
Temperature
(°F - 32°)/1.8 = °C
(1.8 × °C) + 32° = °F
Power
(BHP) 746 = W
(BHP) 0.746 = kW
(W)/746 = BHP
(kW)/0.746 = BHP
102 | Fans and Systems
Annex B. Dual Fan Systems - Series and
Parallel
It is sometimes necessary to install two or more fans
in systems that require higher pressures or airflow
than would be attainable with a single fan. Two fans
may offer a space, cost, or control advantage over a
single larger fan, or it may be simply a field
modification of an existing system to boost pressure
or airflow.
B.1 Fans operating in series
To obtain a system pressure boost, fans are often
installed in series. The fans may be mounted as close
as the outlet of one fan directly attached to the inlet
of the next fan, or they may be placed in remote
locations with considerable distance between fans.
The fans must handle the same mass airflow,
assuming no loss or gains between stages.
The combined total pressure will then be the sum of
each fan’s total pressure (Figure B.1). The velocity
pressure corresponds to the air velocity at the outlet
of the last fan stage. The static pressure for the
combination is the total pressure minus the velocity
pressure and is not the sum of the individual fan
static pressures.
In practice there is some reduction in airflow due to
the increased air density in the later fan stage(s).
There can also be significant loss of airflow due to
non-uniform airflow into the inlet of the next fan.
Sometimes multiple impellers are assembled in a
single housing and this assembly is known as a
“multi-stage” fan. This combination is seldom used in
conventional ventilating and air conditioning systems
but it is not uncommon in special industrial systems.
It is advisable to request the fan manufacturer to
review the proposed system design and make some
estimate of its installed performance.
These types of systems normally have common inlet
and outlet sections, or they may have individual ducts
of equal resistance that join together at equal
velocities. In either case, the characteristic curve is
the sum of the separate airflows for a given static or
total pressure (Figure B.2).
The total performance of the multiple fans will be less
than the theoretical sum if inlet conditions are
restricted or the airflow into the inlets is not straight
(see Section 9.6). Also, adding a parallel fan to an
existing system without modifying the resistance
(larger ducts, etc.) will result in lower than anticipated
airflow due to increased system resistance.
Fans that have a “positive” slope in the pressurevolume curve to the left of the peak pressure curve,
typical of some axial and forward curved centrifugal
fans (see Figure 4.2), can experience unstable
operation under certain conditions. If fans are
operated in parallel in the region of this “positive”
slope, multiple operating conditions may occur.
Figure B.2 illustrates the combined pressure-volume
curve of two such fans operating in parallel.
The closed loop to the left of the peak pressure point
is the result of plotting all the possible combinations
of volume airflow at each pressure. If the system
curve intersects the combined volume-pressure
curve in the area enclosed by the loop, more than
one point of operation is possible. This may cause
one of the fans to handle more of the air and could
cause a motor overload if the fans are individually
driven. This unbalanced airflow condition tends to
reverse readily with the result that the fans will
intermittently load and unload. This "pulsing" often
generates noise and vibration and may cause
damage to the fans, ductwork or driving motors.
Aileron controls in forward curved fan outlets or
dampers near the inlets or outlets may be used to
correct unbalanced airflow or to eliminate pulsations
or reversing operation (See Figure B.3).
B.2 Fans operating in parallel
Suppliers of air handling equipment and designers of
custom systems commonly incorporate two identical,
in parallel fans to deliver large volumes of air while
taking advantage of the space savings offered by
using two smaller fans.
Fans and Systems | 103
PERCENT OF FAN STATIC PRESSURE
SYSTEM
RESISTANCE
200%
SERIES FAN
COMBINED
PRESSURE
CURVE
100%
SINGLE FAN
PRESSURE
CURVE
100%
PERCENT OF FAN AIRFLOW
Figure B.1 - Typical Characteristic Curve of Two Fans Operating in Series
104 | Fans and Systems
STA
BL
ES
YS
TE
M
UN
STA
BL
ES
YS
TE
M
PERCENT OF FAN STATIC PRESSURE
FAN OPERATION NOT
RECOMMENDED IN THIS
RANGE
100
PARALLEL FANS - FAN PRESSURE AT
COMBINED VOLUME
SINGLE FAN PRESSURE
CURVE
200
PERCENT OF FAN AIRFLOW
Figure B.2 - Parallel Fan Operation
AILERON
Figure B.3 - Aileron Control
Fans and Systems | 105
Annex C. Definitions and Terminology
C.1 The air
C.1.1 Air velocity. The velocity of an air stream is its
rate of motion, expressed in m/s (fpm). The velocity
at a plane (Vx) is the average velocity throughout the
entire area of the plane.
C.1.2 Airflow. The airflow at a plane (Qx) is the rate
of airflow, expressed in m3/s (cfm) and is the product
of the average velocity at the plane and the area of
the plane.
C.1.3 Barometric pressure. Barometric pressure
(pb) is the absolute pressure exerted by the
atmosphere at a location of measurement (per AMCA
99-0066).
C.1.4 Pressure-static. Static pressure is the portion
of the air pressure that exists by virtue of the degree
of compression only. If expressed as gauge pressure,
it may be negative or positive (per AMCA 99-0066).
Static pressure at a specific plane (Psx) is the
arithmetic average of the gauge static pressures as
measured at specific points in the traverse of the
plane.
C.1.5 Pressure-velocity. Velocity pressure is that
portion of the air pressure which exists by virtue of
the rate of motion only. It is always positive (per
AMCA 99-0066).
Velocity pressure at a specific plane (Pvx) is the
square of the arithmetic average of the square roots
of the velocity pressures as measured at specific
points in the traverse plane.
C.1.6 Pressure-total. Total pressure is the air
pressure that exists by virtue of the degree of
compression and the rate of motion. It is the
algebraic sum of the velocity pressure and the static
pressure at a point. Thus if the air is at rest, the total
pressure will equal the static pressure (per AMCA 990066).
Total pressure at a specific plane (Ptx) is the algebraic
sum of the static pressure and the velocity pressure
at that plane.
C.1.7 Standard air density. A density of 1.2 kg/m3
(0.075 lbm/ft3) corresponding approximately to air at
20°C (68°F), 101.325 kPa (29.92 in. Hg) and 50%
relative humidity (per AMCA 99-0066).
106 | Fans and Systems
C.1.8 Temperature. The dry-bulb temperature (td) is
the air temperature measured by a dry temperature
sensor. Temperatures relating to air density are
usually referenced to the fan inlet.
The wet-bulb temperature (tw) is the temperature
measured by a temperature sensor covered by a
water-moistened wick and exposed to air in motion.
Readings shall be taken only under conditions that
assure an air velocity of 3.6 to 10.2 m/s (700 to 2000
ft/min) over the wet-bulb and only after sufficient time
has elapsed for evaporative equilibrium to be
attained.
Wet bulb depression is the difference between drybulb and wet-bulb temperatures (td - tw) at the same
location.
C.2 The fan
C.2.1 Blast area. The blast area of a centrifugal fan
is the fan outlet area less the projected area of the
cutoff; see Figure B.6 (per AMCA 99-0066).
C.2.2 Inlet area. The fan inlet area (A1) is the gross
inside area of the fan inlet (see Figure 9.14).
C.2.3 Outlet area. The fan outlet area (A2) is the
gross inside area of the fan outlet.
C.2.4 Fan. (1) A device, which utilizes a power-drive
rotating impeller for moving air or gases. The internal
energy (enthalpy) increase imparted by a fan to a gas
does not exceed 25 kJ/kg (10.75 BTU/lbm). (2) A
device having a power-driven rotating impeller
without a housing for circulating air in a room (per
AMCA 99-0066).
The volume airflow of a fan (Q) is the rate of airflow
in m3/s (cfm) expressed at the fan inlet conditions.
C.2.5 Fan impeller diameter. The fan impeller
diameter is the maximum diameter measured over
the impeller blades.
C.2.6 Fan total pressure. Fan total Pressure (Pt) is
the difference between the total pressure at the fan
outlet and the total pressure at the fan inlet. Pt = Pt1 Pt2 (Algebraic).
Ignoring the losses that exist between the planes of
measurement and the fan, Figures C.1, C.2 and C.3
illustrate fan total pressures for three basic
arrangements for fans connected to external
systems.
Where the fan inlet is open to atmospheric air or
where an inlet bell, as shown in the Figure C.1 is
used to simulate an inlet duct, the total pressure at
the fan inlet (Pt1) is considered to be the same as the
total pressure in the region near the inlet (Pta) where
no energy has been imparted to the air. This is the
location of "still air". The following equations apply:
Where the fan outlet is open to atmospheric air or
where an outlet duct three diameters or less in length
is used to simulate a fan with an outlet duct and the
outlet duct is open to atmospheric air, the total
pressure at the fan outlet is equal to the fan velocity
pressure (Pv). The following equations apply:
Pt = Pt2 - Pt1
Pt2 = Pv
Pt = Pv - Pt1
Pta = 0
Pt = Pt2 - Pt1
Pt1 = Pta = 0
Pt = Pt2
PLANE 1
PLANE 2
Pt2
Pt = Pt2
Figure C.1 - Fan Total Pressure for Installation Type B: Free Inlet, Ducted Outlet
Fans and Systems | 107
PLANE 1
PLANE 2
Pt1
Pt = Pv2 - Pt1
Figure C.2 - Fan Total Pressure for Installation Type C: Ducted Inlet, Free Outlet
PLANE 1
Pt1
PLANE 2
Pt
Pt = Pt2 - Pt1
Figure C.3 - Fan Total Pressure for Installation Type D: Ducted Inlet, Ducted Outlet
108 | Fans and Systems
Pt2
C.2.7 Fan velocity pressure. Fan velocity pressure
(Pv) is the pressure corresponding to the average air
velocity at the fan outlet. Pv = Pv2
Assuming no change in air density or area between
the plane of measurement and the fan outlet, Figure
C.4 illustrates fan velocity pressure.
C.2.8 Fan static pressure. The difference between
the fan total pressure and the fan velocity pressure.
Therefore, fan static pressure is the difference
between the static pressure at a fan outlet and the
total pressure at a fan inlet (per AMCA 99-0066).
Ps = Pt - Pv
Ignoring losses between the planes of measurement
and the fan, Figure C.5 illustrates the fan static
pressure for a fan with ducted inlet and outlet.
Ps = Ps2 - Ps1 - Pv1 (Algebraic)
Where the fan inlet is open to atmospheric air, (free
inlet, ducted outlet), the fan static pressure (Ps) is
equal to the static pressure at the fan outlet.
Ps = Ps2
PLANE 1
Pv = Pv2
Where the fan outlet is open to atmospheric air
(ducted inlet, free outlet), ignoring the SEF, the fan
static pressure (Ps) is equal to the inlet static
pressure (Ps1) less the inlet velocity pressure (Pv1).
Ps = -Ps1 - Pv1
Ps = -(-Ps1) - Pv1
Ps = Ps1 - Pv1
C.3 The system
C.3.1 Equivalent duct diameter. The diameter of a
circle having the same area as another geometric
shape. For a rectangular cross-section duct with
width (a) and height (b), the equivalent diameter is:
(4ab/π)0.5 (per AMCA 99-0066).
C.3.2 Fan performance. Fan performance is a
statement of the volume airflow, static or total
pressure, speed and power input at a stated inlet
density and may include total and static efficiencies.
C.3.3 Fan performance curve. Of the many forms of
fan performance curves, generally all convey
information sufficient to determine fan performance
as defined above. In this manual, ‘fan performance
curve’ refers to the constant speed performance
PLANE 2
Pv2
Figure C.4 - Fan Velocity Pressure for Installation Type B: Free Inlet, Ducted Outlet
Fans and Systems | 109
curve. This is a graphical representation of static or
total pressure and power input over a range of
volume airflow at a stated inlet density and fan
speed. It may include static or total efficiency curves.
The range of volume airflow that is covered generally
extends from shutoff (zero airflow) to free delivery
(zero fan static pressure). The pressure curves that
appear are generally referred to as the pressurevolume curves.
C.3.4 Normalized fan curve. A normalized fan curve
is a constant speed curve in which the fan
performance values appear as percentages, with
100% airflow at free delivery, 100% fan static
pressure at shutoff, and 100% power at the maximum
power input point.
C.3.5 Point of duty. Point of duty is a statement of
air volume flow rate and static or total pressure at a
stated density and is used to specify the point on
the system curve at which a fan is to operate.
C.3.6 Point of operation. The relative position on a
fan or air curtain performance curve corresponding to
a particular airflow, pressure, power and efficiency
(per AMCA 99-0066).
PLANE 1
Ps1
C.3.7 Point of rating. The specified fan operating
point on its characteristic curve (per AMCA 99-0066).
C.3.8 System. A series of ducts, conduits, elbows,
branch piping, etc., designed to guide the flow of air,
gas or vapor to and from one or more locations.
A fan provides the necessary energy to overcome
the resistance to flow of the system and causes air or
gas to flow through the system. Some components of
a typical system are louvers, grills, diffusers, filters,
heating and cooling coils, air pollution control
devices, burner assemblies, sound attenuators, the
ductwork and related fittings.
C.3.9 System curve. A graphic representation of the
pressure versus volume airflow characteristics of a
particular system.
C.3.10 System Effect Factor (SEF). A pressure loss,
which recognizes the effect of fan inlet restrictions,
fan outlet restrictions, or other conditions influencing
fan performance when installed in the system (per
AMCA 99-0066).
PLANE 2
Pv1
Ps2
Ps = Ps2 - Ps1 - Pv1 (algebraic)
Figure C.5 - Fan Static Pressure for Installation Type D: Ducted Inlet, Ducted Outlet
110 | Fans and Systems
HOUSING
DIVERTER
CU
TO
FF
CENTER PLATE
BLAST AREA
DISCHARGE
OUTLET AREA
SIDE SHEET
BACKPLATE
FF
BLADE
TO
CU
INLET
SCROLL
IMPELLER
FRAME
RIM
BEARING
SUPPORT
INLET COLLAR
Figure C.6 - Terminology for Centrifugal Fan Components
Fans and Systems | 111
CASING
BACKPLATE
RIM
HUB
MOTOR
GUIDE VANE
INLET
BLADE
IMPELLER
INLET BELL
Figure C.7A - Tubular Centrifugal FanDirect Drive
CASING
BLADE
DIFFUSER
HUB
MOTOR
IMPELLER
CASING
Figure C.7B - Tubeaxial Fan-Direct Drive
(Impeller Downstream)
BEARING CASING
BELT TUBE
BLADE
HUB
GUIDE VANE
IMPELLER
Figure C.7C - Vaneaxial Fan-Belt Drive
Figure C.7 - Terminology for Axial and Tubular Centrifugal Fans
112 | Fans and Systems
The Ps required at the fan outlet (C) will be equal to
the pressure drop at the desired airflow. Since there
are no inlet obstructions and the duct near the fan
outlet is the same as used in the test setup, the
published fan performance can be used with no
additional system effect factors applied.
Annex D. Examples of the Convertibility
of Energy from Velocity Pressure to
Static Pressure
SI CONVERSION was done using 249 Pa = 1 in. wg,
1 m3/s = 2118 cfm, 1m/s = .00508 ft/min
D.1 Example of fan (tested with free inlet,
ducted outlet) applied to a duct system
The overall friction of the duct system results in a 747
Pa (3.0 in. wg) pressure drop at an airflow of 1.42
m3/s (3000 cfm).
SI
I-P
A
Free inlet
0.00 Pa
(no SEF)
0.0 in. wg
B-C
Outlet with straight
duct attached for two
or more diameters.
0.00 Pa
(no SEF)
0.0 in. wg
Duct friction at Q =
1.42 m3/s (3000 cfm).
747.00 Pa
(duct design)
3.0 in. wg
C-D
REQUIRED FAN Ps
747.00 Pa
3.0 in. wg
Select a fan for Q = 1.42 m3/s (3000 cfm) and Ps = 747 Pa (3.0 in. wg).
Use manufacturer's data for rpm (N) and power (H).
NO OBSTRUCTION AT FAN INLET
Pv = 124 Pa (0.5 in.wg)
FRICTION 747 Pa (3.0 in.wg)
AT 1.42 m3/s (3000 cfm)
(I-P) in.wg
(SI) Pa
996
4
747
3
Pt
498
2
249
1
0
0
Pv
Ps
A
B C
ATMOSPHERIC PRESSURE
124 Pa
(0.5 in.wg)
D
Figure D.1 - Pressure Gradients - Fan as Tested
Fans and Systems | 113
D.2 Example of fan (tested with free inlet,
ducted outlet), connected to a duct system
and then a plenum
This example includes the same duct system as
described in Example C.1. However, there is a short
outlet duct on the fan followed by a plenum chamber
with cross-sectional area more than 10 times larger
than the area of the duct.
The velocity in the duct from E to F is 14.4 m/s (2830
fpm), equal to a velocity pressure of 124.5 Pa (0.5 in.
wg). At point "F" the Pv is 124.5 Pa (0.5 in. wg), the
Ps is 0.0 Pa (0.0 in. wg), and the Pt is 124.5 Pa (0.5
in. wg). The friction of duct will cause a gradual
increase in Ps and Pt back to point E. If the duct has
a uniform cross-sectional area the Pv will be constant
through this part of the system.
Since there is an energy loss of 49.8 Pa (0.2 in. wg)
as a result of the abrupt contraction from the plenum
to the duct, the Pt requirement in the plenum is
871.15 Pa (3.5 in. wg), Pt at duct entrance = 49.8 Pa
(0.2 in. wg) in contraction loss, or 921.3 Pa (3.7 in.
wg) Pt.
Air flowing across the plenum from D to E will have a
relatively low velocity and the Pv in the plenum will be
0.0 Pa (0.0 in. wg) since the velocity is negligible.
At point D, there is an abrupt expansion energy loss
equal to the entire Pv in the duct discharging into the
plenum. The outlet duct between the fan and the
plenum is 2.5 equivalent diameters long. It is the
same as used during the fan rating test. The Ps in the
outlet duct (also the Ps in the plenum) is the same as
the Ps as measured during the rating test.
This example requires a fan to be selected for 921.30
Pa (3.7 in. wg) at 1.42 m3/s (3000 cfm). Compare this
with the previous selection of 747 Pa (3.0 in. wg) Ps
at 1.42 m3/s (3000 cfm).
SI
I-P
C-D
Outlet duct on fan as tested
0.00 Pa
D
Pv loss (also Pt loss) as
result of air velocity decrease.
Ps does not change from
duct to plenum at D.
0.00 Pa
Contraction loss - plenum
to duct
49.80 Pa
(part of duct system)
0.2 in. wg
Ps energy required to
create velocity at E
124.50 Pa
(part of duct system)
0.5 in. wg
Duct friction at Q =
1.42 m3/s (3000 cfm)
747.00 Pa
3.0 in. wg
921.30 Pa
3.7 in. wg
E
E
E-F
REQUIRED FAN Ps
(no SEF)
0.0 in. wg
Solution:
Select a fan for Q = 1.42 m3/s (3000 cfm) and Ps = 921.30 Pa (3.7 in. wg)
Use manufacturer's data for rpm (N) and power (H).
114 | Fans and Systems
0.0 in. wg
2.5 DIA.
NEGLIGIBLE
LOSS
Pv = 124 Pa (0.5in.wg)
FRICTION 747 Pa (3.0 in.wg)
AT 1.42 m3/s (3000 cfm)
(I-P) in.wg
(SI) Pa
1046 Pa (3.7 in.wg)
1245
5
996
4
747
3
498
2
922 Pa (3.7 in.wg)
922 Pa
(3.7 in.wg)
Pt
747 Pa (3.0 in.wg)
249
1
0
0
A
B C
D
E
Pv
124 Pa
(0.5 in.wg)
Ps
ATMOSPHERIC PRESSURE
F
Figure D.2 - Pressure Gradients - Plenum Effect
Fans and Systems | 115
D.3 Example of fan with free inlet, free outlet
- fan discharges directly into plenum and
then to duct system (abrupt expansion at fan
outlet)
energy is lost. In these applications, the energy loss
and the System Effect Factor may exceed the fan
outlet velocity pressure as defined in terms of "fan
outlet area".
This example is similar to the plenum effect example
except the duct at the fan outlet has been omitted.
The fan discharges directly into the plenum.
The SEF for fans without outlet duct was obtained as
follows:
GIVEN:
It may seem unreasonable that the System Effect
loss at the fan outlet is greater than the defined fan
outlet velocity. Fans with cutoffs must generate
higher velocities at the cutoff plane (blast area) than
in the outlet duct (outlet area). This higher velocity (at
cutoff) is partially converted to Ps when outlet ducts
are used as on fan tests. When fans with cutoffs are
"bulk-headed" into plenums or discharge directly into
the atmosphere as with exhausters, all the velocity
Fan
Blast Area
= 0 .6
Outlet Area
Fan outlet velocity = 14.4 m/s
(2830 fpm) No outlet duct
System Effect Curve = R-S, (from Figure 8.3)
SEF = 149.4 Pa (0.6 in. wg), (from Figure 7.1) at 14.4
m/s (2830 fpm) velocity and system curve R)
SI
I-P
149.40 Pa
0.6 in. wg
0.0 in. wg
B-C
SEF
(see above)
B-C
Pv loss (also Pt loss) as
result of air velocity decrease.
Ps does not change from
duct to plenum at C
0.00 Pa
contraction loss - plenum
to duct
49.80 Pa
(part of duct system)
0.2 in. wg
Ps energy required to
create velocity at D
124.50 Pa
(part of duct system)
0.5 in. wg
duct friction at Q =
1.42 m3/s (3000 cfm)
747.00 Pa
(duct design)
3.0 in. wg
D
D
D-E
REQUIRED FAN Ps
1070.70 Pa
Solution:
Select a fan for 1.42 m3/s (3000 cfm) Q and 1070.70 Pa (4.3 in. wg) Ps.
Use manufacturer's data for rpm (N) and power (H).
116 | Fans and Systems
4.3 in. wg
Pv = 124 Pa (0.5 in.wg)
(I-P) in.wg
FRICTION 747 Pa (3.0 in.wg)
AT 1.42 m3/s (3000 cfm)
149 Pa
(0.6 in.wg)
SEF
(SI) Pa
922 Pa (3.7 in.wg)
1245
5
996
4
747
3
498
2
872 Pa (3.5 in.wg)
Pt
747 Pa (3.0 in.wg)
249
1
0
0
A
B C
D
Pv
124 Pa
(0.5 in.wg)
Ps
E
ATMOSPHERIC PRESSURE
Figure D.3 - Pressure Gradients - Abrupt Expansion at Fan Outlet
Fans and Systems | 117
Three SEFs are shown in this example:
D.4 Example of fan used to exhaust with
obstruction in inlet, inlet elbow, inlet duct,
free outlet
1) System Effect Curve R (see Figure 9.5 for a 3
piece inlet elbow with R/D ratio of 1 and no duct
between the elbow and the fan inlet).
This example is an exhaust system. Note the entry
loss at point A. An inlet bell will reduce this loss.
2) System Effect Curve U (see Figure 9.14 for a
bearing in the fan inlet which obstructs 10% of the
inlet).
On the suction side of the fan, Ps will be negative, but
Pv is always positive.
3) System Effect Curve R (from Figure 8.3 for a fan
discharging to atmosphere with no outlet duct).
Fan Pv = 124.5 Pa (0.5 in. wg)
SI
A
I-P
Entrance loss - sharp
edge duct
99.60 Pa
(duct design)
0.4 in. wg
A-B
Duct friction at 1.42 m3/s (3000 cfm)
747.00 Pa
(duct design)
3.0 in. wg
B
SEF 1
149.40 Pa
0.6 in. wg
C
SEF 2
49.80 Pa
0.2 in. wg
E
Fan Pv
124.50 Pa
0.5 in. wg
E
SEF 3
149.40.Pa
0.6 in. wg
REQUIRED FAN Pt
1319.70 Pa
5.3 in. wg
Fan Ps = fan Pt - fan Pv
Fan Ps (SI) = 1319.70 Pa – 124.5 Pa = 1195.2 Pa
Fan Ps (I-P) = 5.3 in. wg - 0.5 in. wg = 4.8 in wg
Solution:
Select a fan for 1.42 m3/s (3000 cfm) Q and 1195.2 Pa (4.8 in. wg) Ps
Use manufacturer's data for rpm (N) and power (H).
118 | Fans and Systems
ABRUPT
DISCHARGE SEF
149 Pa (0.6 in.wg)
Pv = 124 Pa (0.5 in.wg)
FRICTION 747 Pa (3.0 in.wg)
AT 1.42 m3/s (3000 cfm)
(I-P) in.wg
ELBOW SEF
149 Pa (0.6 in.wg)
(SI) Pa
OBSTRUCTION SEF
50 Pa (0.2 in.wg)
+249 +1
ATMOSPHERIC PRESSURE
0
0
-249
-1
-498
-2
-747
-3
100 Pa (0.4 in.wg)
149 Pa (0.6 in.wg)
REQUIRED
Pt
Pv
-847 Pa (-3.4 in.wg)
Ps
-996 Pa (4.0 in.wg)
224 Pa (0.9 in.wg)
-996
-4
-1245 -5
-1171 Pa (4.7 in.wg)
-971 Pa (3.9 in.wg)
A
B
C
D
E
-1121 Pa (4.5 in.wg)
FAN INLET
Figure D.4 - Pressure Gradients - Exhaust System
Fans and Systems | 119
Annex E. References
These references contain additional information related to the subject of this manual:
1. ANSI/AMCA 210-99, Laboratory Methods of Testing Fans for Aerodynamic Performance Rating, Air Movement
and Control Association International, Inc., 30 West University Drive, Arlington Heights, IL, 60004-1893 U.S.A.,
1999.
2. AMCA Publication 200-95, Air Systems, Air Movement and Control Association International, Inc., 30 West
University Drive, Arlington Heights, IL, 60004-1893 U.S.A., 1995.
3. AMCA Publication 202-98, Troubleshooting, Air Movement and Control Association International, Inc., 30 West
University Drive, Arlington Heights, IL, 60004-1893 U.S.A., 1997.
4. ASHRAE Handbook, HVAC Systems and Equipment, 1996, The American Society of Heating, Refrigerating
and Air Conditioning Engineers, Inc., 1791 Tullie Circle N.E., Atlanta, GA, 30329 U.S.A., 1996, (Chapter 18
Fans).
5. Traver, D. G., System Effects on Centrifugal Fan Performance, ASHRAE Symposium Bulletin, Fan Application,
Testing and Selection, The American Society of Heating, Refrigerating and Air Conditioning Engineers, Inc.,
1791 Tullie Circle N.E., Atlanta, GA, 30329 U.S.A., 1971.
6. Christie, D. H., Fan Performance as Affected By Inlet Conditions, ASHRAE Transactions, Vol. 77, The
American Society of Heating, Refrigerating and Air Conditioning Engineers, Inc., 1791 Tullie Circle N.E.,
Atlanta, GA, 30329 U.S.A., 1971.
7. Zaleski, R. H., System Effect Factors For Axial Flow Fans, AMCA Paper 2011-88, AMCA Engineering
Conference, Air Movement and Control Association International, Inc., 30 West University Drive, Arlington
Heights, IL, 60004-1893 U.S.A., 1988.
8. Roslyng, O., Installation Effect on Axial Flow Fan Caused Swirl and Non-Uniform Velocity Distribution,
Institution of Mechanical Engineers (IMechE), 1 Birdcage Walk, London SW1H 9JJ, England, 1984.
9. Clarke, M. S., Barnhart, J. T., Bubsey, F. J., Neitzel, E., The Effects of System Connections on Fan
Performance, ASHRAE RP-139 Report, The American Society of Heating, Refrigerating and Air Conditioning
Engineers, Inc., 1791 Tullie Circle N.E., Atlanta, GA, 30329 U.S.A., 1978.
10. Madhaven, S., Wright, T., J. DiRe, Centrifugal Fan Performance With Distorted Inflows, The American Society
of Mechanical Engineers, 345 East 47th Street, New, York, NY, 10017 U.S.A., 1983.
11. Cory, W. T. W., Fan System Effects Including Swirl and Yaw, AMCA Paper 1832-84-A5, AMCA Engineering
Conference, Air Movement and Control Association International, Inc., 30 West University Drive, Arlington
Heights, IL, 60004-1893 U.S.A., 1984.
12. Cory, W. T. W., Fan Performance Testing and Effects of the System, AMCA Paper 1228-82-A5, AMCA
Engineering Conference, Air Movement and Control Association International, Inc., 30 West University Drive,
Arlington Heights, IL, 60004-1893 U.S.A., 1984.
13. Galbraith,
L.E.,
Discharge
Diffuser
Effect
on
Performance
Axial Fans, AMCA Paper 1950-86-A6, AMCA Engineering Conference, Air Movement and Control Association
International, Inc., 30 West University Drive, Arlington Heights, IL, 60004-1893 U.S.A., 1986.
14. Industrial Ventilation –23rd Edition, American Conference of Governmental Industrial Hygienists, 1330 Kemper
Meadow Drive, Cincinnati, OH 45240-1634 U.S.A., 1998.
15. Fans and Systems, John E. Thompson and C. Jack Trickler, The New York Blower Company, Chemical
Engineering, March 21, 1983, pp. 48-63
16. AABC National Standards, Chapter 8, Volume Measurements, Associated Air Balance Council, 1518 K Street
NW, Suite 503, Washington, DC 20005 U.S.A.
120 | Fans and Systems
Troubleshooting
202
1. Introduction
After the installation of an air moving system is completed, a system sometimes fails to achieve its designed
performance.
This part of the AMCA Fan Application Manual will help you identify what is wrong and decide how to correct it.
2. Procedure for Troubleshooting
2.1 Look in the "Master Troubleshooting Appendices" for a subject which corresponds with the apparent problem.
2.2 Check each of the "Probable Causes" listed.
2.3 If the cause of the trouble is not found proceed through the "System Checklist" (see Section 4).
2.4 If the problem has still not been solved, it is now advisable to contact the representative of the fan manufacturer.
He should be given the results of the "System Checklist" and the additional information listed in Section 5.1.
2.5 The fan manufacturer or his representative will analyze the information submitted, as outlined in Section 5.2.
With this information and, if necessary, an on-site inspection, he may be able to explain why the system is not
achieving its design performance and may recommend changes in the system or the fan installation which will
overcome the problem.
3. Safety Precautions
Before checking the fan and system it will be necessary to shut down the fan. During inspection the fan must be
electrically isolated and all disconnect switches and other controls LOCKED in the "OFF" position. Where these
are in locations remote from the fan, prominent DO NOT START signs should also be in place.
CAUTION - Even when LOCKED out electrically, fans located outdoors or in a parallel or series fan system may
be subject to "wind-milling." Therefore, as an added precaution, the impeller should be secured to physically restrict
rotational movement.
4. System Checklist
Poor system performance may arise from a number of causes including:
•
•
•
•
•
•
•
Improper installation or assembly of the fan
Damage in handling or transit
System design error
Deterioration of the system
Faulty controls
Poor fan selection
A combination of several factors.
A systematic check of the items listed here should help identify the problem - or problems - and allow suitable
corrective action to be taken.
SYSTEM CHECKLIST
A) While the impeller is coasting to a stop, see if it is rotating in the proper direction (see Figures 4.1, 4.2 and
4.3).
B) Make certain the impeller rotation is correct for the housing (guide vanes of vaneaxial and tubular centrifugal
fans) and not installed backwards (see Figures 4.1, 4.2, and 4.3).
Note: Tubeaxial fan rotation is the same as shown in Figure 4.2 except without the guide vanes.
ROTATION
RADIAL
BLADE
BACKWARD
INCLINED
AIRFOIL
RADIAL
TIP
BACKWARD
CURVED
FORWARD
CURVED
Figure 4.1 - Types of Centrifugal Fan Impellers
Note: Fan manufacturers describe the rotation of the fan impeller as being "clockwise" or "counterclockwise”.
•
For AXIAL fans when viewing the INLET or OUTLET as specified by the Fan Manufacturer
•
For CENTRIFUGAL fans when viewing the DRIVE SIDE (see AMCA Standard 99-2406)
•
For TUBULAR CENTRIFUGAL fans when viewing the OUTLET (see AMCA Standard 99-2410).
122 | Troubleshooting
ROTATION
ROTATION
INLET
GUIDE VANES
OUTLET
GUIDE VANES
AIRFLOW
AIRFLOW
Figure 4.2 - Vaneaxial Fan Impellers and Guide Vanes
ROTATION
GUIDE
VANES
AIRFLOW
Figure 4.3 - Tubular Centrifugal Impeller
Troubleshooting | 123
C) If the fan is belt driven:
1) Are the drive pulley (Motor Sheave) and the driven pulley (Fan Sheave) in alignment? Improper alignment
of the sheaves can cause excessive power consumption (high amperage), squealing belts, shortened belt
and sheave life and high axial vibration.
FAN
FAN
MOTOR
IMPROPER SHEAVE ALIGNMENT
MOTOR
PROPER SHEAVE ALIGNMENT
Figure 4.4 - Sheave Alignment
2) Are the belts loose? Loose belts that flap or slip can cause excessive noise and vibration. Slipping belts
will also affect fan speed and reduce belt and sheave life.
Belts should be tensioned to the belt manufacturer's recommendations. Tension of the drive belts should be
adjusted for stretching after the first 48 hours of operation. Caution! Excessive belt tension will reduce fan
and motor bearing life.
3) Are the belts and/or sheaves worn? If so, an immediate replacement could save down time at a later
date. Where more than one belt is used, replace with a new set of matched belts.
IMPROPER BELT TENSION
PROPER BELT TENSION
(SEE BELT MANUFACTURER’S SPECIFICATIONS)
Figure 4.5 - Belt Tension
D) Check the flow surfaces (passages between the inlets, impeller blades and inside of housing) for
cleanliness. A 2 mm (0.0625 in.) buildup of dirt on the flow surface could impair fan performance and/or
cause vibration.
E) Are there any gouges, tears, holes, erosion or corrosion in the impeller blades, rims or backplate: inlets
and/or housing? If so, report the approximate size and location to the fan manufacturer and discontinue
operation until repairs are made.
F) Is any foreign matter trapped in the impeller, housing or ductwork (loose insulation, papers, ice, etc.)? If so,
remove.
124 | Troubleshooting
G) Are coils, heaters, filters, ducts, etc. dirt laden? If so, clean or replace. Remove any non-essential
obstructions to airflow in elbows, shutters, transformations, dampers, bird-screens, etc. Verify that all
dampers (control, backdraft, fire, etc.) are adjusted to the proper settings.
H) Have all the parts supplied with the fan been installed?
I)
Are there any obstructions to airflow near the fan inlets? Objects such as pipes, airflow measurement
stations, ductwork, columns, belt guards, belt drives, etc. could adversely affect the output of the fan. For
more information, see AMCA Publication 201 Fans and Systems.
J)
Are the fan outlet connections correctly designed and installed? Duct takeoffs, or obstructions in the fan
outlet could adversely affect the output of the fan (see AMCA Publication 201).
K) See Figure 4.6A for typical centrifugal fan inlet-impeller relationships and Figure 4.6A for typical axial fan
housing-impeller relationships. A few simple measurements as indicated on these figures can tell the
manufacturer if a problem exists in this area. Note: several measurements should be taken around the
entire inlet/housing circumference to determine the average, maximum and minimum values.
I.S. = INLET SPACING
I.G. = INLET GAP
I.G.
R.C.
R.C. = RUNNING CLEARANCE
(INLET CENTERED ON IMPELLER)
I.S.
I.S.
I.S.
I.S. SHOULD BE SAME BOTH SIDES
FOR DOUBLE WIDTH FAN
Figure 4.6A - Typical Centrifugal Fan Inlet-Impeller Relationships
R.C.
C.
R.C. = RUNNING CLEARANCE
(IMPELLER CENTERED IN HOUSING)
C. = VANE CLEARANCE
Figure 4.6B - Typical Axial Fan Housing-Impeller Relationships
Troubleshooting | 125
L)
On a double-width fan, is the approach to both inlets identical? Airflow should be symmetrical about the
centerline of the fan housing (see Figure 4.7A). Non-symmetrical airflow can lead to decreased air
performance. Belt drives, belt guards and motors can cause non-symmetrical airflow to the inlets if too
severely restricted (see AMCA Publication 201).
L
L
P
AIRFLOW
P
Figure 4.7A - Symmetrical Inlet Airflow
L
L NOT EQUAL TO M
N NOT EQUAL TO R
M
N
AIRFLOW
R
Figure 4.7B - Non-Symmetrical Inlet Airflow
Figure 4.8A - Turning Vanes in Elbow Adjacent to Centrifugal Fan
M) Are turning vanes installed in elbows that are too close to the fan inlet or discharge (see Figures 4.8A and
4.8B). Published pressure losses through elbows are based on a uniform velocity profile. Turning vanes
help achieve this uniform flow (see AMCA Publication 201).
126 | Troubleshooting
Figure 4.8B - Turning Vanes in Elbow Adjacent to Axial Fan
N) If the fan is equipped with an inlet vane damper, check the operation as follows:
1)
Do not rely on the control arm position alone for locating the position of the inlet vane damper blades
without first checking visually to see that the inlet vane damper position agrees with the position of the
control arm.
2)
If the unit is a double-width fan equipped with inlet vanes or inlet vane damper control, both inlet vane
dampers must be synchronized (the inlet vane dampers must be in the same relative position with
respect to the impeller on both inlets). If the inlet vane dampers are not synchronized, there will be
unbalanced airflow between inlets resulting in deficient air performance, unbalanced thrust on the
bearings and/or a surge condition in the fan.
3)
Make certain that inlet vane dampers are of the proper rotation with respect to the impeller.
As the vanes close they should cause the entering air to spin in the same direction as the impeller.
4) Are the inlet vane dampers correctly positioned for the design operating conditions? If not, the desired
pressure-volume of the fan will not be realized (see Figure 4.9).
100% Open
Percent of No Delivery Pressure
100
80
75% Open
50% Open
60
25% Open
40
20
0
20
40
60
80
100
Percent of Free Delivery Volume
Figure 4.9 - Typical Pressure-Volume Curve for Operation with Inlet Vane Damper Control
Troubleshooting | 127
O) Inspect any ductwork or plenums approaching the fan inlets for the possibility of inducing swirl of air into the
inlet. Pre-swirl of air entering the fan inlet can reduce the fan performance (see Figures 4.10A and 4.10B).
(See AMCA Publication 201.)
IMPELLER
ROTATION
IMPELLER
ROTATION
AIR SWIRL IN SAME DIRECTION
AS IMPELLER ROTATION
AIR SWIRL IN OPPOSITE DIRECTION
TO IMPELLER ROTATION
Figure 4.10A - Pre-Rotation
Figure 4.10B - Counter-Rotation
P) After completing the above steps and making sure the fan and system are safe to start, remove all DO NOT
START signs on disconnect switches and override systems and put the unit back into operation.
Q) Inspect the entire system including the fan, fan plenum and all ductwork for significant air leaks.
Air leaks may be detected by sound, smoke, feel, soapy solution, etc. Some common air leak sources are:
access doors, coils, duct seams, fan outlet connection, etc. Significant air leaks must be sealed.
5. Fan Manufacturer's Analysis
If the cause of the trouble has still not been found after completing the "System Checklist," the fan manufacturer
should be consulted.
The fan manufacturer will review the information provided concerning the system and recommend an appropriate
course of action.
5.1 Data required for analysis
To make a complete analysis of the problem, in addition to the results of the "System Checklist," the manufacturer
will need:
A)
Complete plans (drawings) including all ductwork, location, size, model and manufacturer of all fans,
motors, coils, dampers, etc. with all pertinent dimensions for the complete system as actually installed.
B)
If the fan/air handling system fails to achieve the design performance, the measured performance and the
design performance figures should be supplied.
C)
A copy of the system design calculations.
D)
A copy of the specifications and any addendums.
128 | Troubleshooting
E)
If a separate air performance test has been conducted on the installed fan, a statement of measured fan
performance along with a copy of the test data, the type of test and instrumentation, and the measurement
location of the airflow rate and pressure determinations should be supplied. A statement of fan performance
should contain:
1) Fan total pressure (Pt) rise or fan static pressure (Ps)
2) Airflow rate (Q)
3) Power (H)
4) Fan speed (N)
5) Air density (ρ)
5.2 Probable manufacturer action
Among other actions, the fan manufacturer will:
A)
Assess the probable accuracy of the field performance measurements (see AMCA Publication 203 Field
Performance Measurements of Fan Systems).
B)
Examine the system drawings (plans) for any System Effect losses (see AMCA Publication 201) which were
not allowed for in the system design calculations or the original fan selection.
C)
Reassess the fan performance, accounting for System Effect losses established in step B to the designed
fan/system performance (see Figure 5.1).
D)
Check whether the fan selection is correct for the application.
SYSTEM EFFECT
DUCT SYSTEM CURVE
DESIGN FAN OPERATING POINT
DESIGN PRESSURE
FAN CATALOG
PRESSURE-VOLUME CURVE
FAN OPERATING POINT
W/SYSTEM EFFECT
THEORETICAL PRESSURE-VOLUME
CURVE ACCOUNTING FOR SYSTEM EFFECT
LOSSES (FIELD PERFORMANCE MEASUREMENTS
TO BE COMPARED AGAINST THIS CURVE).
DESIGN VOLUME
Figure 5.1
Troubleshooting | 129
The information obtained through the checklists in this manual should help in defining the necessary corrective
action.
In most cases, if the troubleshooting procedure has been followed carefully and impartially it will be apparent
whether the system has been built and installed in accordance with the design drawings, whether the fan was
properly selected and suitable allowances made for the appropriate System Effect Factors, or whether the fan is
not performing up to its published ratings.
6. Conclusion
By intelligent application of the procedures outlined in this manual it should be possible to find the cause of a
performance problem in any air moving system.
Identification of a problem associated directly with the fan may require the assistance of the fan manufacturer.
Recognition of the cause of the trouble will usually be a major step toward curing it. Corrective measures may
include alterations to the system, modifications to the fan outlet or inlet connections, adjustments to the fan, etc.
In many cases an increase in the fan speed may be decided upon but it is extremely important that the fan shall
not be operated above its cataloged maximum speed or the maximum speed recommended by the
manufacturer. Excessive speed may result in catastrophic impeller failure. If a speed increase proves to be
an acceptable alternative, then the motor should also be checked for its capacity to handle the increased fan power.
130 | Troubleshooting
MASTER TROUBLESHOOTING APPENDICES
Annex A. Noise
SOURCE
PROBABLE CAUSE
A-1
IMPELLER HITTING INLET OR HOUSING . . . . . . . . . . . . . a.
b.
c.
d.
e.
f.
g.
h.
Impeller not centered in inlet or housing.
Inlet or housing damage.
Crooked or damaged impeller.
Shaft loose in bearing.
Impeller loose on shaft.
Bearing loose in bearing support.
Bent shaft.
Misaligned shaft and bearings.
A-2
IMPELLER HITTING CUTOFF . . . . . . . . . . . . . . . . . . . . . . . a.
b.
c.
d.
Cutoff not secure in housing.
Cutoff damaged.
Cutoff improperly positioned.
Impeller improperly positioned.
A-3
DRIVE . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. Sheave not tight on shaft (motor or fan).
b. Belts hitting belt tube or belt guard.
c. Belts too loose. Adjust for belt
stretching after 48 hours of operation.
d. Belts too tight.
e. Belts wrong cross-section.
f. Belts not "Matched" in length on
multi-belt drive.
g. Variable pitch sheaves not
adjusted so each groove has same
pitch diameter (multi-belt drive).
h. Misaligned sheaves.
i. Belts worn.
j. Motor, motor base or fan not
securely anchored.
k. Belts oily or dirty.
l. Improper drive selection.
m. Loose key.
A-4
COUPLING . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. Coupling unbalanced, misaligned, loose
or may need lubricant.
b. Loose key.
A-5
BEARING . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
c.
d.
e.
f.
g.
h.
i.
A-6
SHAFT SEAL SQUEAL . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
c.
d.
Defective bearing.
Needs lubrication.
Loose on bearing support.
Loose on shaft.
Seals misaligned.
Foreign material inside bearing.
Worn bearing.
Fretting corrosion between inner race
and shaft.
Bearing not sitting on flat surface.
Needs lubrication.
Misaligned.
Bent shaft.
Bearing loose on support.
Troubleshooting | 131
A-7
IMPELLER . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. Loose on shaft.
b. Defective impeller. DO NOT
OPERATE FAN. LOCK OUT
THE FAN ELECTRICALLY AND
CONTACT THE MANUFACTURER.
c. Unbalance.
d. Coating loose.
e. Worn as result of abrasive or corrosive
material moving through airflow
passages.
f. Blades rotating close to structural
member.
g. The number of blades may coincide
with an equal number of structural
members.
h. Loose key.
A-8
HOUSING . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. Foreign material in housing.
b. Cutoff or other housing part loose
(rattling during operation).
A-9
MOTOR . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
c.
d.
e.
f.
g.
h.
A-10
SHAFT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. Bent
b. Undersized. May cause noise at
impeller, bearings or sheave.
A-11
HIGH AIR VELOCITY . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. Ductwork too small for application.
b. The installed fan may be too small for
the application.
c. Registers or grilles too small for
application.
d. Heating or cooling coil with
insufficient face area for application.
OBSTRUCTION IN HIGH VELOCITY
AIR STREAM MAY CAUSE RATTLE, OR
PURE TONE WHISTLE . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. Dampers.
b. Registers.
c. Grilles.
d. Sharp elbows.
e. Sudden expansion in ductwork.
f. Sudden contraction in ductwork.
g. Turning vanes.
A-12
A-13
Lead-in cable not secure.
AC hum in motor or relay.
Starting relay chatter.
Noisy motor bearings.
Single phasing a three phase motor.
Low voltage.
Cooling fan striking shroud.
Poor motor/inverter match, more
noticeable at low speeds.
PULSATION OR SURGE . . . . . . . . . . . . . . . . . . . . . . . . . . . a. Restricted system causes fan to
operate at poor point of rating.
b. Fan too large for application.
c. Ducts vibrate at same frequency as fan
pulsations.
d. Rotating stall.
132 | Troubleshooting
e. Inlet vortex surge.
f. Distorted inlet airflow.
A-14
AIR LEAKS AND/OR OBSTRUCTED FLOW . . . . . . . . . . . . a. Air leaks in ductwork.
1) Bad joint connections;
2) Holes or tears;
3) Obstructions inside duct.
b. Fins on coils.
c. Registers or grilles.
A-15
RATTLES AND/OR RUMBLES . . . . . . . . . . . . . . . . . . . . . . . a. Vibrating ductwork.
b. Vibrating cabinet parts.
c. Vibrating parts not isolated from
building.
Troubleshooting | 133
Annex B. Insufficient Air Flow
SOURCE
B-1
PROBABLE CAUSE
FAN . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
c.
d.
e.
f.
g.
h.
i.
j.
B-2
DUCT SYSTEM . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
c.
d.
e.
Impeller installed backwards.
Impeller running backwards.
Impeller blade angle setting.
Cutoff missing or improperly installed.
Impeller not centered with inlet collar(s).
Fan rpm below design:
1) Incorrect sheave or sheave setting;
2) Incorrect motor rpm;
3) Low voltage to motor;
4) Speed controller set too low;
CAUTION! DO NOT INCREASE FAN
SPEED BEYOND THE FAN
MANUFACTURER'S
RECOMMENDATIONS. ALSO, WHEN
INCREASING FAN SPEED, MONITOR
MOTOR AMPS SO AS NOT TO
EXCEED MOTOR NAMEPLATE AMPS.
Impeller/inlet dirty or clogged.
Improper running clearance.
Improper inlet cone to impeller fit.
Improperly set inlet vane or damper.
Actual system is more restrictive (more
resistance to airflow) than expected.
Dampers closed.
Registers closed.
Air leaks in supply ducts.
Insulating duct liner loose.
B-3
FILTERS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
Dirty or clogged.
Replacement filter with greater than
specified pressure drop.
B-4
COILS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
Dirty or clogged.
Incorrect fin spacing.
B-5
RECIRCULATION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .a.
Internal cabinet air leaks in bulkhead
separating fan outlet (pressure zone)
from inlet(s) (suction zone).
Air leaks around fan outlet at
connection through cabinet bulk-head.
Elbows, cabinet walls or other
obstructions restrict airflow. Inlet
obstructions cause more restrictive
systems but do not cause increased
negative pressure readings near the fan
inlet(s) (see System Effects in AMCA
Publication 201). Fan speed may be
increased to counteract the effect of
restricted fan inlet(s).
CAUTION! DO NOT INCREASE FAN
SPEED BEYOND THE FAN MANUFACTURER'S RECOMMENDATIONS.
ALSO, WHEN INCREASING FAN
b.
B-6
OBSTRUCTED FAN INLETS . . . . . . . . . . . . . . . . . . . . . . . .a.
134 | Troubleshooting
SPEED, MONITOR MOTOR AMPS SO
AS NOT TO EXCEED MOTOR
NAMEPLATE AMPS.
B-7
NO STRAIGHT DUCT AT FAN OUTLET . . . . . . . . . . . . . . . a.
Fans which are normally used in duct
system are tested with a length of
straight duct at the fan outlet. If there is
no straight duct at the fan outlet,
decreased performance may result. If it
is not practical to install a straight
section of duct at the fan outlet, the fan
speed may need to be increased to
overcome this pressure loss (see
System Effects in AMCA Publication
201).
CAUTION! DO NOT INCREASE FAN
SPEED BEYOND THE FAN MANUFACTURER'S RECOMMENDATIONS.
ALSO, WHEN INCREASING FAN
SPEED, MONITOR MOTOR AMPS SO
AS NOT TO EXCEED MOTOR
NAMEPLATE AMPS.
B-8
OBSTRUCTION IN HIGH VELOCITY AIR STREAM . . . . . . a.
b.
c.
d.
Obstruction near fan outlet or inlet(s).
Sharp elbows near fan outlet or inlet(s).
Improperly designed turning vanes.
Projections, dampers or other
obstruction in a part of the system
where air velocity is high.
Troubleshooting | 135
Annex C. Airflow High - (Too Much Airflow)
SOURCE
PROBABLE CAUSE
C-1
SYSTEM . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
c.
d.
e.
f.
Oversized ductwork.
Access door open.
Registers or grilles not installed.
Dampers set to by-pass coils.
Filter(s) not in place.
System resistance low.
C-2
FAN . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
Fan speed too fast.
Improper impeller blade angle.
136 | Troubleshooting
Annex D. Static Pressure Wrong
SOURCE
D-1
PROBABLE CAUSE
SYSTEM, FAN OR INTERPRETATION
OF MEASUREMENT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . GENERAL DISCUSSION:
The velocity pressure at any point of
measurement is a function of the velocity of
the air and its density.
The static pressure at a point of
measurement in the system is a function of
the system design (resistance to airflow), air
density and the amount of air flowing through
the system.
The static pressure measured in a "loose" or
oversized system will be less than the static
pressure in a "tight" or undersized system for
the same airflow rate.
In most systems, pressure measurements
are indicators of how the installation is
operating. These measurements are the
result of airflow and as such are useful
indicators in defining system characteristics.
Field static pressure measurements rarely
correspond with laboratory static pressure
measurements unless the fan inlet(s) and
fan outlet conditions of the installation are
exactly the same as the inlet(s) and outlet
conditions in the laboratory.
Also see D-2 through D-6, E-2, F-1 and G-1
for specific cases.
Static Pressure Low, Airflow Correct
D-2
GAS DENSITY . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Pressures will be less with high
temperature gases or at high altitudes.
Static Pressure Low, Airflow High
D-3
SYSTEM . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . System has less resistance to airflow
than expected. This is a common
occurrence. Fan speed may be reduced to
obtain desired airflow rate. This will reduce
power consumption (operating cost).
Static Pressure Low, Airflow Low
D-4
SYSTEM . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
Fan inlet(s) and/or outlet conditions not
same as tested. See System Effect
Factors, in AMCA Publication 201.
Troubleshooting | 137
D-5
FAN . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
c.
d.
e.
f.
g.
h.
i.
j.
Impeller installed backwards.
Impeller running backwards.
Improper impeller blade angle.
Cutoff missing or improperly installed.
Impeller not centered with inlet collar(s).
Fan speed too slow.
Impeller/inlet dirty or clogged.
Improper running clearance.
Improper inlet cone to impeller fit.
Improperly set inlet vanes or damper.
Static Pressure High, Airflow Low
D-6
DUCT SYSTEM . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
c.
d.
Actual system is more restrictive (more
resistance to airflow) than designed.
Dampers closed.
Registers closed.
Insulating duct liner loose.
D-7
FILTERS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
Dirty or clogged.
Replacement filter with greater than
specified pressure drop.
D-8
COILS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
Dirty or clogged.
Fin spacing too close.
138 | Troubleshooting
Annex E. Power High
SOURCE
E-1
PROBABLE CAUSE
FAN . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
c.
d.
E-2
SYSTEM . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
c.
d.
Backward Inclined impeller installed
backwards.
Fan speed too high.
Forward Curve or Radial Blade
impeller operating below design
pressures.
Incorrect impeller blade angle.
Oversized ductwork.
Face and bypass dampers
oriented so coil dampers are open at
same time bypass dampers are open.
Filter(s) left out.
Access door open.
Note: The causes listed under E-2 pertain
primarily to radial blade, radial tip and
forward curve centrifugal fans, i.e., fans that
exhibit rising power curves. Normally,
backward inclined, backward curve or
backward inclined airfoil centrifugal fans and
axial flow fans do not fall into this category.
E-3
GAS DENSITY . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .a.
Calculated power requirements based
on light gas (e.g. high temperature) but
actual gas is heavy (e.g. cold start-up).
E-4
FAN SELECTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .a.
Fan not selected at efficient point of
rating.
Troubleshooting | 139
Annex F. Fan Does Not Operate
SOURCE
F-1
PROBABLE CAUSE
ELECTRICAL OR MECHANICAL . . . . . . . . . . . . . . . . . . . . . Mechanical and electrical problems are
usually straight-forward and are normally
analyzed in a routine manner by service
personnel. In this category are such items:
a.
b.
c.
d.
e.
f.
g.
h.
i.
j.
140 | Troubleshooting
Blown fuses.
Broken belts.
Loose pulleys.
Electricity turned off.
Impeller touching housing.
Wrong voltage.
Motor too small and overload protector
has broken circuit.
Low voltage, excessive line drop or
inadequate wire size.
Load inertia too large for motor.
Seized bearing.
Annex G. Premature Failure
SOURCE
PROBABLE CAUSE
G-1
BELTS, BEARINGS, SHEAVES IMPELLERS
HUBS, ETC.GENERAL DISCUSSION . . . . . . . . . . . . . . . . . Each fan component is designed to operate
satisfactorily for a reasonable lifetime. Fans
intended for heavy duty service are made
especially for that type of service. For
example, Class I fans are intended for
operation below certain limits of pressure
and outlet velocity. Class II fans are
designed for higher operating limits (see
AMCA Standard 99-2408). Not all
components are limited by the same factors,
e.g. limiting factors may be power, RPM,
temperature, impeller tip speed, torque,
corrosive atmospheres, expected life, etc.
Also see A-2, A-5, A-6, A-7.
G-2
COUPLINGS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
Coupling unbalanced, misaligned, loose
or may need lubricant.
Loose key.
G-3
SHAFT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
Bent.
Undersized.
G-4
BEARINGS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
c.
d.
e.
f.
Over-lubrication or under- lubrication.
Locking collar or set screw loose.
Excessive belt tension.
False brinelling.
Wrong lubricant.
Undersized shaft.
G-5
DRIVES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
c.
d.
e.
Over-tensioned or under-tensioned belts.
Mismatched belts on multi-belt drive.
Misalignment of motor and fan sheaves.
Excessive start-stop cycles.
Set screw on sheave loose.
Troubleshooting | 141
Annex H. Vibration
AERODYNAMIC VIBRATION (PULSATION)
SOURCE
H-1
PROBABLE CAUSE
AIRSTREAM . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
c.
d.
e.
f.
Operating fan left of peak in unstable
(stall) region.
Poor inlet conditions which generate air
turbulence.
1. Partially obstructed inlet(s).
2. Sharp elbow at fan inlet(s) and/or
outlet.
System pulsation which is transmitted to
the fan.
Blade Pass frequency -- number of
blades x fan RPM.
Guide vane frequency -- number of
vanes x fan RPM.
Fan support frequency -- motor or
bearing supports, belt tube, on axial
fans.
MECHANICAL VIBRATION
H-2
UNBALANCED IMPELLERS . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
c.
d.
e.
f.
g.
h.
i.
j.
k.
l.
H-3
DRIVE OR COUPLING . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
c.
d.
e.
f.
g.
H-4
LOOSE FASTENERS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
c.
d.
e.
f.
142 | Troubleshooting
Material build-up on impeller.
Impeller components subject to
abrasion, corrosion or impact.
Moisture inside hollow airfoil blades.
Lost balance weight.
Excessive temperature.
Impeller (blades) not tracking evenly.
Eccentricity:
1. Bore off center.
2. Bore on angle (see H-2.f)
Improper key-to-keyway length.
Impeller rubbing stationary equipment.
Shaft seal rub.
Inverter drives.
Motor torque pulses (on single phase
motors).
Alignment:
1. Improper assembly.
2. Shift during handling or shipment.
Worn, loose, or mismatched belts.
Eccentric sheaves or couplings.
Belt resonance.
Worn or chipped sheaves.
Improper key-to-keyway length.
Unbalanced sheave(s) or coupling.
Impeller set screws.
Bearing set screws.
Drive component set screws.
Fan mounting bolts.
Bearing bolts.
Motor bolts.
H-5
FAN SUPPORT STRUCTURE . . . . . . . . . . . . . . . . . . . . . . . a.
b.
c.
d.
e.
f.
g.
h.
Support structure natural frequencies.
Insufficient mass or stiffness.
Large unsupported mass.
Lack of lateral support, particularly with
fans mounted on isolators.
Fan skewed (bolted down to uneven
surface).
Broken support members.
Short-circuited or improperly adjusted
isolators.
Fan (mounted on isolators) with rigid
inlet(s) and outlet connections.
H-6
SHAFT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
Bent.
Undersized.
H-7
BEARINGS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
c.
Misalignment.
Worn out; loose.
Too much, too little, or incorrect
lubricant.
H-8
BACKGROUND EXCITATION
(FLANKING TRANSMISSION) . . . . . . . . . . . . . . . . . . . . . . .a.
b.
c.
d.
Interconnected piping.
Heavy machinery transmitting vibration
through foundation.
Fan(s) mounted on floating roofs.
Wind-loading of fan mounted on
isolators.
ELECTRICALLY INDUCED VIBRATION
H-9
ELECTRIC MOTOR . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.
b.
c.
d.
e.
f.
Loose stator laminations.
Broken rotor bar.
Stator problems.
Phase unbalance.
Eccentric rotor.
Capacitor-start motor causes "shudder"
during coast-down.
Troubleshooting | 143
Field Performance
Measurements of Fan Systems
203
1. Introduction
2. Scope
Performance ratings of fans are developed from
laboratory tests made according to specified
procedures on standardized test setups. In North
America, the standard is ANSI/AMCA Standard 210 /
ANSI/ASHRAE 51 Laboratory Methods of Testing
Fans for Rating.
The recommendations and examples in this
publication may be applied to all types of centrifugal,
axial, and mixed flow fans in ducted or nonducted
installations used for heating, ventilating, air
conditioning, mechanical draft, industrial process,
exhaust, conveying, drying, air cleaning, dust
collection, etc. Although the word air is used when
reference is made in the general sense to the
medium being handled by the fan, gases other than
air are included in the scope of this publication.
In actual systems in the field, very few fans are
installed in conditions reproducing those specified in
the laboratory standard. This means that, in
assessing the performance of the installed fansystem, consideration must be given to the effect on
the fan’s performance of the system connections,
including elbows, obstructions in the path of the
airflow, sudden changes of area, etc. The effects of
system conditions on fan performance is discussed in
Section 5, and more completely in AMCA Publication
201, Fans and Systems.
A major problem of testing in the field is the difficulty
of finding suitable locations for making accurate
measurements of flow rate and pressure. Sections
9.3 and 10.3 outline the requirements of suitable
measurement sections.
Because these problems and others will require
special consideration on each installation, it is not
practical to write one standard procedure for the
measurement of the performance of all fan-systems
in the field. This publication offers guidelines to
making performance measurements in the field
which are practical and flexible enough to be applied
to a wide range of fan and system combinations.
Because of the wide variety of fan types and systems
encountered in the field, Annex A includes examples
of a number of different field tests. In most cases,
these examples are based on actual tests which have
been conducted in the field.
Before performing any field test, it is strongly
recommended that the following AMCA publications
be carefully reviewed:
AMCA Publication 200 - Air Systems
AMCA Publication 201 - Fans and Systems
AMCA Publication 202 - Troubleshooting
AMCA Standard 210 - Laboratory Methods of Testing
Fans for Rating
Measurement of sound, vibration, and stress levels
are not within the scope of this publication.
3. Types of Field Tests
There are three general categories of field tests:
A) General
Fan
System
Evaluation
A measurement of the fan-system’s performance
to use as the basis of modification or adjustment
of the system.
B) Acceptance Test -A test specified in the sales
agreement to verify that the fan is achieving the
specified performance.
C) Proof of Performance Test -A test in response
to a complaint to demonstrate that the fan is
meeting the specified performance requirement.
As acceptance and proof of performance tests are
related to contract provisions, they are usually
subject to more stringent requirements and are
usually more costly than a general evaluation test. In
the case of large fans used in industrial applications
and of mechanical draft fans used in the electrical
power generation industry the performance of a field
test may be part of the purchase agreement between
the fan manufacturer and the customer. In addition to
Publication 203, AMCA Standard 803 Site
Performance Test Standard-Power Plant and
Industrial Fans defines the conditions which must be
met to achieve higher accuracy of measurement. In
new installations of this type, it is desirable to include
a suitable measuring section in the design.
Agreement must be reached on the test method to be
used prior to performance of the test.
4. Alternatives to Field Tests
In some cases, considerations such as cost and
problems of making accurate measurements may
make the following alternative methods of testing
worth investigation:
A) Testing the fan before installation in a laboratory
equipped to perform tests in accordance with
AMCA Standard 210. Limitations in laboratory
test facilities may preclude tests on full size fans.
In this case, the full size fan can be tested at the
installation site in accordance with AMCA
Standard 210. This will usually require the
installation of special ductwork.
B) Testing a reduced scale model of the fan in
accordance with AMCA Standard 210 and
determining the performance of the full size fan
as described in AMCA Publication 802, Power
Plant Fans – Establishing Performance Using
Laboratory Methods.
C) Testing a reduced scale model of the complete
fan and system using the test methods outlined
in this publication.
Tests conducted in accordance with AMCA Standard
210 will verify the performance characteristics of the
fan but will not take into account the effect of the
system connections on the fan’s performance (see
Section 5).
5. System Effect Factors
AMCA Publication 201, Fans and Systems, deals in
detail with the effect of system connections on fan
performance. It gives system effect factors for a wide
variety of obstructions and configurations which may
affect a fan’s performance.
System Effect Factor (SEF) is a pressure loss which
recognizes the effect of fan inlet restrictions, fan
outlet restrictions, or other conditions influencing fan
performance when installed in the system.
SYSTEM EFFECT FACTORS (SEFs) ARE
INTENDED TO BE USED IN CONJUNCTION WITH
THE SYSTEM RESISTANCE CHARACTERISTICS
IN THE FAN SELECTION PROCESS. Where SEFs
are not applied in the fan selection process, SEFs
must be applied in the calculations of the results of
field tests. This is done for the purpose of allowing
direct comparison of the test results to the design
static pressure calculation. Thus, for a field test, the
fan static pressure is defined as:
Ps = Ps2 - Ps1 – Pv1 + SEF 1 + SEF 2 + + SEF n
146 | Field Performance Measurement
Examples of the application of SEFs in determining
the results of field tests are included in Annex A.
In field tests of fan-system installations in which
system effects have not been accounted for, it is
important that their sources be recognized and their
magnitudes be established prior to testing.
The alternative to dealing with a large magnitude
SEF is to eliminate its source. This requires revisions
to the system. This alternative course of action is
recommended when swirl exists at the fan inlet (see
Publication 201, Figure 9.8). The effect on fan
performance as a result of swirl at the inlet is
impossible to estimate accurately as the system
effect is dependent upon the degree of swirl. The
effect can range from a minor amount to an amount
that results in the fan-system performance being
completely unacceptable.
6. Fan Performance
Fan performance is a statement of fan flow rate, fan
total or static pressures, and fan power input at stated
fan speed and fan air density. Fan total or static
efficiencies may be included. The fan air density is
the density at the fan inlet. The fan flow rate is the
volume flow rate at the fan inlet density.
7. Referenced Planes
Certain locations within a fan-system installation are
significant to field tests. These locations are
designated as follows:
Plane 1: Plane of fan inlet
Plane 2: Plane of fan outlet
Plane 3: Plane of Pitot-static tube traverse for
purposes of determining flow rate
Plane 4: Plane of static pressure measurement
upstream of fan
Plane 5: Plane of static pressure measurement
downstream of fan
The use of the numerical designations as subscripts
indicate that the values pertain to those locations.
8. Symbols and Subscripts
SYMBOL
A
D
De
FLA
H
HL
Hmo
kW
L
N
NLA
NPH
NPV
Ps
Psx
Pt
Ptx
Pv
Pvx
pb
pe
pp
px
Q
Qi
Qx
SEF
T
td
tw
V
∆Px,x’
∆Ps
ρ
ρx
Σ
DESCRIPTION
UNIT
9.1 General
Area of cross-section
Diameter
Equivalent diameter
Full load amps
Fan power input
Power transmission loss
Motor power output
Electrical power
Length
Speed of rotation
No load amps
Nameplated horsepower
Nameplated volts
Fan static pressure
Static pressure at Plane x
Fan total pressure
Total pressure at Plane x
Fan velocity pressure
Velocity pressure at Plane x
Barometric pressure
Saturated vapor pressure at tw
Partial vapor pressure
Absolute pressure at Plane x
Fan flow rate
Interpolated flow rate
Flow rate at Plane x
System effect factor
Torque
Dry-bulb temperature
Wet-bulb temperature
Velocity
Pressure loss between
Planes x and x’
Pressure loss across damper
Fan gas density
Gas density at Plane x
Summation sign
ft2
ft
ft
amps
hp
hp
hp
kilowatts
ft
rpm
amps
hp
volts
in. wg
in. wg
in. wg
in. wg
in. wg
in. wg
in. Hg
in. Hg
in. Hg
in. Hg
cfm
cfm
cfm
in. wg
lb-in.
°F
°F
fpm
Determine fan flow rate using the area, velocity
pressure, and density at the traverse plane and the
density at the fan inlet. The velocity pressure at the
traverse plane is the root mean square of the velocity
pressure measurements made in a traverse of the
plane. The flow rate at the traverse plane is
calculated by converting the velocity pressure to its
equivalent velocity and multiplying by the area of the
traverse plane.
Airflow direction
---
SUBSCRIPT
c
r
x
1
2
3
4
5
9. Fan Flow Rate
in. wg
in. wg
lbm/ft3
lbm/ft3
---
DESCRIPTION
Value converted to specified conditions
Reading
Plane 1, 2, 3, ..., as appropriate
Plane 1 (fan inlet)
Plane 2 (fan outlet)
Plane 3 (plane of Pitot-static traverse for
purpose of determining flow rate
Plane 4 (plane of static pressure
measurement upstream of fan)
Plane 5 (plane of static pressure
measurement downstream of fan)
9.2 Velocity measuring instruments
Use a Pitot-static tube of the proportions shown in
Annex B or a double reverse tube, shown in Annex C,
and an inclined manometer to measure velocity
pressure. The velocity pressure at a point in a gas
stream is numerically equal to the total pressure
diminished by the static pressure. The Pitot-static
tube is connected to the inclined manometer as
shown in Annex F. The double reverse tube is
connected to the inclined manometer as shown in
Annex C.
9.2.1 Pitot-static tube. The Pitot-static tube is
considered to be a primary instrument and need not
be calibrated if maintained in the specified condition.
It is suited for use in relatively clean gases. It may be
used in gases that contain moderate levels of
particulate matter such as dust, water, or dirt,
provided certain precautions are employed (see
Section 15).
9.2.2 Double reverse tube. The double reverse tube
is used when the amount of particulate matter in the
gas stream impairs the function of the Pitot-static
tube. The double reverse tube requires calibration. It
is important that the double reverse tube be used in
the same orientation as used during calibration. Mark
the double reverse tube to indicate the direction of
the gas flow used in its calibration.
9.2.3 Inclined manometers. Inclined manometers
are available in both fixed and adjustable range
types. Both types require calibration. The adjustable
range type is convenient in that it may be adjusted at
the test site to the range appropriate to the velocity
pressures which are to be measured. It is adjusted by
changing the slope to any of the various fixed
settings and by changing the range scale
accordingly. Each setting provides a different ratio of
the length of the indicating column to its indicated
height. Adjustable range type manometers in which
the slope may be fixed at 1:1, 20:1, and intermediate
ratios are available (see Figure 10 in Annex G).
Field Performance Measurement | 147
The accuracy of the manometer used in the
measurement of velocity pressures is of prime
importance. Select a manometer that will provide an
acceptable degree of accuracy; consider the range,
slope, quality, scale graduations, indicating fluid of
the instrument and the range of the velocity
pressures to be measured. The graph in Annex G
indicates the effect of expected resolution of
manometer readings on the accuracy of velocity
determinations. The basis for this graph is described
in Section 9.6. Determine velocities in the very low
range more accurately by using a manometer with a
slope of 20:1. Due to practical limitations in length, its
use is restricted to measurements where the
velocities are very low. Also, errors in velocity
determinations made by using a Pitot-static tube and
manometer exceed normally acceptable values at
velocity pressure readings less than 0.023 in. wg.
This corresponds to a velocity of approximately 600
fpm for air of 0.075 lbm/ft3 density.
9.2.4 Low velocity instruments. Normally, velocities
encountered in the field test situations are well in
excess of 600 fpm. Therefore, recommendations
regarding
alternate
test
procedures
and
instrumentation for use for velocities less than 600
fpm are not presented in this publication.
Descriptions of various types of instruments used to
determine range velocities are presented in Annex J.
Most of the instruments require frequent calibration,
and some are not suited for use in high temperature,
dirty, wet, corrosive, or explosive atmospheres. If it is
necessary to use one of these instruments, the
procedure for its use, its calibration, and the expected
accuracy of results should be agreed upon by all
interested parties.
9.3 Location of traverse plane
For field tests, suitable test measurement station
locations must be provided in the system. When
suitable locations are not available, consider making
temporary or permanent alterations to the ducting for
improved test accuracy.
For free inlet, free outlet fans, convert a free inlet,
free outlet fan to a ducted inlet, free outlet fan by the
addition of a temporary duct. Estimate free inlet, free
outlet fan flow rate by measuring other parameters
and interpreting certified ratings performance (see
Section 17.1).
than 75% of the velocity pressure measurements
are greater than 1/10 of the maximum
measurement (see Figure 9.1)
2) The flow streams should be at right angles to the
traverse plane. Variations from this flow condition
as a result of swirl or other mass turbulence are
considered acceptable when the angle between
the flow stream and the traverse plane is within
10 degrees of a right angle. The angle of the flow
stream in any specific location is indicated by the
orientation of the nose of the Pitot-static tube that
produces the maximum velocity pressure reading
at the location.
3) The cross-sectional shape of the airway in which
the traverse plane is located should not be
irregular. Proper distribution of traverse points
and accurate determination of the area of the
traverse plane are difficult to achieve when the
airway does not conform closely to a regular
shape.
4) The cross-sectional shape and area of the airway
should be uniform throughout the length of the
airway in the vicinity of the traverse plane. When
the divergence or convergence of the airway is
irregular or more than moderate in degree,
significantly nonuniform flow conditions may
exist.
5) The traverse plane should be located to minimize
the effects of gas leaks between the traverse
plane and the fan.
6) When it is necessary to locate the traverse plane
in a converging or diverging airway (not
recommended), note that the traverse plane and
area is located at the tip of the Pitot-static tube.
A location well downstream in a long, straight run of
uniform cross-section duct will usually provide
acceptable conditions for the Pitot traverse plane.
When locating the traverse plane close to the fan, as
is often done in order to minimize the effect of
leakage, flow conditions upstream of the fan are
usually more suitable. In some installations, more
than one traverse plane may be required in order to
account for the total flow (Annex A contains
examples).
A Pitot traverse plane suitable for the measurements
used to determine flow rate are as follows:
When a field test is anticipated, particularly when the
requirement for a field test is an item in the
specifications, the system designer should provide a
suitable traverse plane location in the system.
1) The velocity distribution should be uniform
throughout the traverse plane. The uniformity of
distribution is considered acceptable when more
When the fan is ducted outlet and the traverse plane
is to be located downstream from the fan, the
148 | Field Performance Measurement
Pv MAX
Pv MAX
10
A: IDEAL Pv DISTRIBUTION
Pv MAX
Pv MAX
10
B: GOOD Pv DISTRIBUTION
(ALSO SATISFACTORY FOR FLOW INTO FAN INLETS.
MAY BE UNSATISFACTORY FOR FLOW INTO INLET
BOXES - MAY PRODUCE SWIRL IN BOXES)
Pv MAX
Pv MAX
10
Pv MAX
Pv MAX
10
60%
80%
C: SATISFACTORY Pv DISTRIBUTION - MORE
THAN 75% OF Pv READINGS GREATER THAN:
Pv MAX
10
(UNSATISFACTORY FOR FLOW INTO FAN INLETS OR
INLET BOXES)
Pv MAX
10
D: DO NOT USE
UNSATISFACTORY Pv DISTRIBUTION LESS THAN 75% OF Pv READINGS GREATER
THAN:
Pv MAX
10
(UNSATISFACTORY FOR FLOW INTO FAN INLETS OR
INLET BOXES)
Pv MAX
Pv MAX
10
Pv MAX
40%
35%
20%
35%
E: DO NOT USE
UNSATISFACTORY Pv DISTRIBUTION
LESS THAN 75% OF Pv READINGS GREATER
THAN:
P MAX
v
F: DO NOT USE
UNSATISFACTORY Pv DISTRIBUTION
LESS THAN 75% OF Pv READINGS GREATER
THAN:
P MAX
v
10
10
(UNSATISFACTORY FOR FLOW INTO FAN INLETS OR
INLET BOXES)
(UNSATISFACTORY FOR FLOW INTO FAN INLETS OR
INLET BOXES)
Figure 9.1 - Typical Velocity Pressure Distributions Encountered in Velocity
Pressure Measurement Planes in Fan-System Installations
Field Performance Measurement | 149
MEASUREMENT PLANE
De
MIN.
2
12 in. MIN.
Y
WHERE: De =
4YZ
π
INLET BOX DAMPERS
Z
Note: The measurement plane should be located a minimum of ½ De from the inlet cone, but not less than 12 in.
from the leaving edge of the damper blades.
Figure 9.2
STACK
VELOCITY
PROFILE
Note: Spiral vortex may form when fan discharges directly into a stack or similar arrangement.
Figure 9.3
150 | Field Performance Measurement
traverse plane should be situated a sufficient
distance downstream from the fan to allow the flow to
diffuse to a more uniform velocity distribution and to
allow the conversion of velocity pressure to static
pressure. Annex P provides guidance for the location
of the traverse plane in these cases. The location of
the traverse plane on the inlet side of the fan should
not be less than ½ equivalent diameter from the fan
inlet. Regions immediately downstream from elbows,
obstructions and abrupt changes in airway area are
not suitable traverse plane locations. Regions where
unacceptable levels of swirl are usually present, such
as the region downstream from an axial flow fan that
is not equipped with straightening vanes, should be
avoided. Swirl may form when a fan discharges
directly into a stack or similar arrangement (see
Figure 9.2).
9.3.1 Inlet box location. When the traverse plane
must be located within an inlet box, the plane should
be located a minimum of 12 inches downstream from
the leaving edges of the damper blades and not less
than ½ equivalent diameter upstream from the edge
of the inlet cone (see Figure 9.3). Do not locate
traverse points in the wake of individual damper
blades. In the case of double inlet fans, traverses
must be conducted in both inlet boxes in order to
determine the total flow rate.
9.3.2 Alternative locations. On occasion, an
undesirable traverse plane location is unavoidable, or
each of a limited number of prospective locations
lacks one or more desirable qualities. In such cases,
the alternatives are:
1) Accept the most suitable location and evaluate
the effects of the undesirable aspects of the
location on the accuracy of the test results. In
some instances, the estimated accuracy may
indicate that the results of the test would be
meaningless, particularly in acceptance tests and
proof of performance tests.
2) Provide a suitable location by modifying the
system. This course of action is recommended
for acceptance tests and proof of performance
tests. The modifications may be temporary,
permanent, minor or extensive, depending on the
specific conditions encountered. When the inlet
side of the fan is not ducted but is designed to
accept a duct, consider installing a short length of
inlet duct to provide a suitable traverse plane
location. This duct should be of a size and shape
to fit the fan inlet, a minimum of 2 equivalent
diameters long and equipped with a bell shaped
or flared fitting at its inlet. The traverse plane
should be located a minimum of ½ equivalent
diameters from the fan inlet and not less than 1½
equivalent diameters from the inlet of the duct.
Where the duct is small, its length may
necessarily be greater than 2 equivalent
diameters in order to ensure that the tip of the
Pitot-static tube is a minimum of 1½ equivalent
diameters from the duct inlet. This short length of
duct should produce no significant addition to the
system resistance, but in some cases it may alter
the pattern of flow into the fan impeller, and
thereby affect the performance of the fan slightly.
9.4 The traverse
Annex H contains recommendations for the number
and distribution of measurement points in the
traverse plane. If the flow conditions at the traverse
plane are less than satisfactory, increase the number
of measurement points in the traverse to improve
accuracy.
Since the flow at a traverse plane is never strictly
steady, the velocity pressure measurements
indicated by the manometer will fluctuate. Each
velocity pressure measurement should be mentally
averaged on a time-weighted basis. Any velocity
pressure measurement that appears as a negative
reading is to be considered a velocity pressure
measurement of zero and included as such in the
calculation of the average velocity pressure.
When it is necessary to locate the traverse plane in a
converging or diverging airway, orient the nose of the
Pitot-static tube such that it coincides with the
anticipated line of the flow stream. This is particularly
important at measurement points near the walls of
the airway (see Annex A-1A).
No appreciable effect on Pitot-static tube readings
occur until the angle of misalignment between the
airflow and the tube exceeds 10 degrees.
9.5 Flow rate calculations
9.5.1 Flow rate at traverse plane. The flow rate at
the traverse plane is calculated as follows:
Q3 = V3A3
Where:
A3 = the area of the traverse plane
V3 = the average velocity at the traverse plane
= 1096 (Pv3/ρ3)0.5
ρ3 = the density at the traverse plane
Pv3 = the root mean square velocity pressure at the
traverse plane
= [∑(Pv3r)0.5 / number of readings]2
Field Performance Measurement | 151
Pv3r is the velocity pressure reading, corrected for
manometer calibration and where applicable,
corrected for the calibration of the double reverse
tube. It is important that the calibration of the double
reverse tube be applied correctly. The use of the
calibration of the double reverse tube is described in
Annex C.
9.5.2 Continuity of mass. The calculations of fan
flow rate are based on considerations of continuity of
mass, and as such, it is assumed that no mass is
added or removed from the gas stream between the
traverse plane and the fan inlet. In the general
application, having determined the flow rate and
density at the traverse plane, the flow rate at any
location, (x), in the fan-system installation may be
calculated, providing the density at this location is
known and the assumption noted above is valid, i.e.:
Qx = Q3 (ρ3/ρx)
9.5.3 Fan flow rate, single traverse plane. Where a
single traverse plane is used, the calculation of the
fan flow rate is:
Q = Q1
= Q3 (ρ3/ρ1)
Where:
Q3 and ρ3 = as described in Section 9.5.1
ρ1 = the density at the fan inlet
9.5.4 Fan flow rate, multiple traverse planes.
When it is necessary to use more than one traverse
plane in order to account for the total flow:
Q = Q1
= Q3a (ρ3a/ρ1) + Q3b (ρ3b/ρ1) + ... + Q3n (ρ3n/ρ1)
9.6 Accuracy
The performance item of major concern in most fansystem installations is the flow rate. Every effort
should be made to improve the accuracy of the flow
rate determination. The uncertainty analysis
presented in Annex T indicates that the uncertainties
in flow rate determinations will range from 2% to
10%. This range is based on considerations of the
conditions that are encountered in most field test
situations. This includes instances in which the
conditions at the Pitot traverse plane do not conform
to all of the qualifications indicated in Section 9.3.
The graph in Annex G provides guidance for
improving the accuracy of the flow rate
152 | Field Performance Measurement
determinations. This graph indicates the effect of
expected resolution of velocity determinations. This
effect is shown for several manometer slope ratios.
For all ratios, the expected resolution used as a basis
for the graph is the length of indicating column
equivalent to 0.05 in. wg in a manometer with slope
ratio of 1:1. As indicated in the graph, reading
resolution uncertainty can be significant. However,
this uncertainty can be controlled by selecting a
manometer with a slope suited to the velocity
pressures to be measured and by avoiding regions of
very low velocity in the selection of the traverse plane
location. Reading resolution uncertainties exceed
normally acceptable values at velocity pressures less
than 0.023 in. wg. This corresponds to a velocity of
approximately 600 fpm for air of 0.075 lbm/ft3 density.
Generally, ducts are sized for velocities considerably
in excess of 600 fpm. Velocities less than 600 fpm
may exist in certain sections of the system in some
installations, but these sections can usually be
avoided. Do no use a Pitot-static tube and
manometer to determine velocities in the low ranges
associated with filters and cooling coils in air
conditioning, heating, and ventilating units. In some
instances, the uncertainties incurred in the
determinations of low velocity flows may be
acceptable. For example, an uncertainty of 15% in
the determination of the flow rate in a branch duct
that accounts for 20% of the total flow rate for the
system affects the accuracy of the total flow rate
determination by only 3%.
In addition to low range velocities, other conditions
may exist at the traverse plane which can
significantly affect the accuracy of the flow rate
determination. These include nonuniform velocity
distribution, swirl, and other mass turbulence.
Improve the accuracy of the flow rate determination
by avoiding these conditions in the selection of the
traverse plane location, or improve the conditions by
modifying the system.
10. Fan Static Pressure
10.1 General
Determine fan static pressure by using the static
pressures at the fan inlet and outlet, the velocity
pressure at the fan inlet, and applicable System
Effect Factors. The use of System Effect Factors in
the determination of fan static pressure is described
in Section 5. The velocity pressure at the fan inlet is
the calculated average velocity pressure at this
location, and as such, its determination is based on
the fan flow rate, the density at the fan inlet, and the
fan inlet area. The static pressures at the fan inlet and
outlet may be obtained directly by making pressure
measurements at these locations; or they may be
determined by making pressure measurements at
other locations, upstream and downstream of the fan.
In the latter case, the determinations must account
for the effects of velocity pressure conversions and
pressure losses, as may occur between the
measurement planes and the planes of interest.
10.2 Pressure measuring instruments
This section describes only the instruments for use in
measuring static pressure. Instruments for use in the
other measurements involved in the determination of
fan static pressure are described in Section 13.
Use a Pitot-static tube of the proportions shown in
Annex B, a double reverse tube as shown in Annex
C, or a side wall pressure tap as shown in Annex E,
and a manometer to measure static pressure.
10.2.1 Pitot-static tube. The comments that appear
in Section 9.2 regarding the use and calibration of the
Pitot-static tube are applicable to its use in the
measurement of static pressures.
10.2.2 Double reverse tube. The double reverse
tube cannot be used to measure static pressure
directly. It must be connected to two manometers and
the static pressure for each point of measurement
must be calculated. Both the manometer connections
and the method of calculation are shown in Annex C.
10.2.3 Pressure tap. The pressure tap does not
require calibration. Use no fewer than four taps
located 90 degrees apart. In rectangular ducts, a
pressure tap should be installed near the center of
each wall. It is important that the inner surfaces of the
duct in the vicinities of the pressure taps be smooth
and free from irregularities, and that the velocity of
the gas stream does not influence the pressure
measurements.
10.2.4 Manometers. A manometer with either
vertical or inclined indicating column may be used to
measure static pressure. Inclined manometers used
to measure static pressures require calibration and
should be selected for the quality, range, slope, scale
graduations, and indicating fluid necessary to
minimize reading resolution errors.
10.3 Static pressure measurements
It is important that all static pressure measurements
be referred to the same atmospheric pressure, and
this atmospheric pressure be that for which the
barometric pressure is determined.
Make static pressure measurements near the fan
inlet and the fan outlet, and where the airway
between the measurement plane and the plane of
interest is straight and without change in crosssectional area. Then the duct friction loss between
the measurement plane and the plane of interest is
usually insignificant, and considerations of velocity
pressure conversions and calculations of pressure
losses for duct fitting and other system components
can be avoided.
When a system component is situated between the
measurement plane and the plane of interest, the
pressure loss of the component must be calculated
and credited to the fan. The calculation of the
pressure loss is usually based on the component’s
performance ratings, which may be obtained from the
manufacturer of the item.
If there is a change in area between the
measurement plane and the plane of interest, then
the calculation of the static pressure at the plane of
interest must account for velocity pressure
conversion and include any associated pressure
loss. When the change in area is moderate and
gradual, the conversion of velocity pressure is
considered to occur without loss and the static
pressure is calculated on the basis of no change in
total pressure between the measurement plane and
the plane of interest. This assumes that the duct
friction loss between the two planes is negligible.
When the change in area is an abrupt and sizable
enlargement, as in a duct leading into a large
plenum, the loss is considered to be equivalent to the
velocity pressure in the smaller area, and the static
pressure at the plane of interest is considered to be
the same as the static pressure at the measurement
plane. This assumes that the velocity pressure in the
larger area and the duct friction loss are negligible.
10.3.1 Location of the measuring plane. When the
fan is ducted outlet, the static pressure measurement
plane downstream of the fan should be situated a
sufficient distance from the fan outlet to allow the flow
to diffuse to a more uniform velocity distribution and
to allow the conversion of velocity pressure to static
pressure. See Annex P for guidance in locating the
measurement plane in these cases. In general,
pressure taps should be used if it is necessary to
measure static pressure in the immediate vicinity of
the fan outlet. The static pressure at this location is
difficult to measure accurately with a Pitot-static tube
due to the existence of turbulence and localized high
velocities. If the surface conditions or the velocities at
the duct walls are unsuited for the use of pressure
taps, then a Pitot-static tube must be used with
extreme care, particularly in aligning the nose of the
tube with the lines of the flow streams.
The location of the static pressure measurement
Field Performance Measurement | 153
plane upstream of the fan should not be less than ½
equivalent diameter from the fan inlet. In the event
that static pressure measurements must be made in
an inlet box, the measurement plane should be
located as indicated in Figure 9.2. In the case of
double inlet fans, static pressure measurements must
be made in both inlet boxes in order to determine the
average static pressure on the inlet side of the fan.
In general, the qualifications for a plane well suited
for the measurement of static pressure are the same
as those for the measurement of velocity pressure,
as indicated in Section 9.3:
negative. By definition, positive values are those
measured as being greater than atmospheric
pressures; negative values are those measured as
being less than atmospheric pressure. In all of the
equations in this publication, the values of static
pressures must be entered with their proper signs
and combined algebraically.
10.4.1 Static pressure at measuring planes. The
static pressure at a plane of measurement (x) is
calculated as follows:
Psx =
1) The velocity distribution should be uniform
throughout the traverse plane.
∑P
sxr
number of readings
Where:
2) The flow streams should be at right angles to the
plane.
3) The cross-sectional shape of the airway in which
the plane is located should not be irregular.
4) The cross-sectional shape and area of the airway
should be uniform throughout the length of the
airway in the vicinity of the plane.
5) The plane should be located such as to minimize
the effects of leaks in the portion of the system
that is located between the plane and the fan.
A long, straight run of duct upstream of the
measurement plane will usually provide acceptable
conditions at the plane. Regions immediately
downstream from elbows, obstructions, and abrupt
changes in airway area are generally unsuitable
locations. Regions where unacceptable levels of
turbulence are present should be avoided.
If in any fan-system installation the prospective
locations for static pressure measurement planes
lack one or more desirable qualities, the alternatives
are to accept the best qualified locations and
evaluate the effects of the undesirable aspects of the
conditions on the accuracy of the test results or
provide suitable locations by modifying the system.
10.3.2 When using a Pitot-static tube or a double
reverse tube to measure static pressure, a number of
measurements must be made throughout the plane.
Use Annex H to determine the number and
distribution of the measurement points. When using
pressure taps, a single measurement at each of the
taps located at the plane is sufficient.
Psxr = the static pressure reading, corrected for
manometer calibration
10.4.2 Static pressure at fan inlet or outlet. The
static pressure at the fan inlet, Ps1, and the static
pressure at the fan outlet, Ps2, may be measured
directly in some cases. In most cases, the static
pressure measurements for use in determining fan
static pressure will not be made directly at the fan
inlet and outlet, but at locations a relatively short
distance upstream from the fan inlet and downstream
from the fan outlet. These static pressure
measurements are designated Ps4 and Ps5,
respectively. Static pressure at the fan inlet, Ps1, is
derived as follows:
Pt4 = Pt1 + ∆P4,1
Where:
Pt4 = the total pressure plane of measurement
Pt1 = the total pressure at the fan inlet
∆P4,1 = the sum of the pressure losses between the
two planes
These losses (∆P) include those attributable to duct
friction, duct fittings, other system components, and
changes in airway area. Although ∆P represents a
loss in all cases, it is considered a positive value as
used in the equations in this publication. By
substitution and rearrangement:
Ps1 = Ps4 + Pv4 - Pv1 - ∆P4,1
Similarly, for static pressure at the fan outlet, Ps2:
10.4 Static pressure calculations
Pt2 = Pt5 + ∆P2,5
Static pressure measurements may be positive or
Ps2 = Ps5 + Pv5 - Pv2 + ∆P2,5
154 | Field Performance Measurement
Where:
10.5 Accuracy
The velocity pressures at the various planes can be
determined from the following general equations for
the velocity pressure at a plane of measurement (x):
The uncertainty analyses in Annex T indicate that the
uncertainties in fan static pressure determinations
are expected range from 2% to 8%. This range is
based on considerations of the conditions expected
to be encountered in most field test situations.
Pvx = Pv3 (A3/Ax)2 (ρ3/ρx)
Or:
Pvx = (Qx/1096Ax)2 ρx
Locate the static pressure measurement planes such
that the pressure losses between the measurement
planes and the planes of interest are insignificant.
This will eliminate the uncertainties involved in the
determination of the pressure losses, and the
equations for Ps1 and Ps2 reduce to the following:
Ps1 = Ps4 + Pv4 - Pv1
Ps2 = Ps5 + Pv5 - Pv2
These equations may be used when changes in area
between the measurement planes and the planes of
interest are moderate and gradual, and the pressure
losses associated with conversions of velocity
pressure to static pressure are negligible.
Improve the accuracy of the fan static pressure
determination by avoiding static pressure
measurement plane locations where turbulence or
other unsteady flow conditions will produce
significant uncertainties in the mental averaging of
pressure readings. Other reading resolution
uncertainties are not as significant in the fan static
pressure determination as in the determination of
flow rate. Generally, static pressure measurements
are much greater in magnitude than velocity pressure
measurements, and the selection of a manometer
that will provide reasonably good accuracy is not
usually a problem.
The uncertainty analyses in Annex T and the
resulting anticipated uncertainty range do not
account for uncertainties that may occur in the
following:
•
Determinations of velocity pressure conversions
occurring between the measurement planes and
the planes of the fan inlet or fan outlet. The area
and density values that are involved in these
determinations are usually obtained without
significant uncertainties. However, pressure
losses associated with velocity pressure
conversions are often difficult to determine
accurately.
•
Determinations of other pressure losses
occurring between the measurement planes and
the fan inlet or fan outlet. This includes pressure
losses in ducts, duct fittings, and other system
components. The calculations of these losses
are based on the assumption of uniform flow
conditions. This assumption may not be valid,
and the calculated pressure loss values may be
significantly inaccurate.
•
Determinations of the values of System Effect
Factors. These determinations are based on
limited information, and as such, are subject to
uncertainty.
If, in addition to the losses being negligible there are
no changes in the areas between the measurement
planes and the respective planes of interest, then the
equations are further reduced to:
Ps1 = Ps4
Ps2 = Ps5
These equations may also be used when the only
losses between the measurement planes and the
planes of interest are those associated with changes
in area that are abrupt and sizable enlargements in
the direction of flow. This assumes that the velocity
pressure in the larger area is negligible.
10.4.3 Fan static pressure. The equation for fan
static pressure is:
Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2 + ... + SEF n
Where:
SEF 1, SEF 2, ... SEF n = System Effect Factors that
account for the various System Effects that are
uncorrected and exist at the time of the field test.
Avoid situations requiring these determinations,
thereby eliminating them as sources for uncertainties.
The uncertainties involved in determining the values
of System Effect Factors can be avoided only by
correcting the causes of the System Effects. This
requires alterations to the system.
Field Performance Measurement | 155
11. Fan Power Input
11.1 General
Fan power input data included as part of the fan
performance ratings are normally defined and limited
to either:
•
power input to the fan shaft
•
the total of the power input to the fan shaft and
the power transmission loss
The losses in fan shaft bearings are included in either
case. Since the results of field tests are usually
compared to the rated performance characteristics of
the fan, field test values of fan power input should be
determined on the same basis as that used in the fan
ratings. For belt driven fans, the rated fan power input
may or may not include belt drive losses. The
information regarding the basis of the rated fan
power input accompanies the rating data or is
otherwise available from the fan manufacturer. In
most instances, when a power transmission loss
occurs, the loss will have to be determined and
subtracted from the motor output in order to obtain
the fan power input.
11.2 Power measurement methods
In view of the fact that accuracy requirements for field
test determinations of fan power input vary
considerably, a number of test methods are
recommended. These methods are intended to
provide economical and practical alternatives for
dealing with various levels of accuracy requirements.
11.2.1 Phase current method. This method for
estimating the power output of three phase motors is
based on the relationship of motor current and motor
power output. The method, described in Annex K,
requires measurements of the phase currents and
voltages supplied to the motor while driving the fan.
Depending on the operating load point of the motor, it
may also involve the measurements of the no load
phase currents.
The phase current method is convenient and
sufficiently accurate for most field tests. In this
method, the closer the actual phase current is to the
motor nameplate value of full load amps, the greater
the accuracy. Since fan motors are normally selected
for operation at or near the full load point, this method
provides a reasonably accurate estimate of the
power output of the fan motor. Determine fan power
input by using the motor power output and, where
applicable, the power transmission loss.
156 | Field Performance Measurement
11.2.2 Typical motor performance data. Typical
motor performance data may be used to determine
fan power input. These data, which are referred to as
typical in that the data and the actual performance of
the motor are expected to correspond closely, can
usually be obtained from the motor manufacturer.
The data provided can be in a variety of forms, but
are sufficient to determine motor power output based
on electrical input measurements. It is important that
the power supplied to the motor during the field test
be consistent with that used as the basis for the
motor performance data. The phase voltage should
be stable and balanced, and the average should be
withing 2% of the voltage indicated in the
performance data.
Depending on the form of the typical motor
performance data, motor power output is determined
by one of the following methods:
1) Given the typical motor performance chart of
watts input versus motor power output at a stated
voltage.
Hmo, is the value in the typical motor performance
data that corresponds to the field test
measurement of watts input to the motor.
2) Given the typical motor performance chart of
watts input versus torque output and speed at a
stated voltage.
Use the field test measurement of watts input
and the corresponding typical motor performance
data values of torque output and speed; the
motor power output is calculated as:
Hmo =
T ×N
63025
3) Given the typical motor performance chart of
watts input versus motor efficiency at a stated
voltage.
Use the field test measurement of watts input
and the corresponding typical motor performance
data value of motor efficiency, the motor power
output is calculated as:
Hmo =
watts input × motor efficiency
746
4) Given the typical motor performance chart of
amps versus power factor and motor efficiency at
a stated voltage.
Use the field test measurements of amps input
and volts, and the typical motor performance
data values of power factor (pf) and motor
efficiency, corresponding to the measured amps
input; the motor power output is calculated as:
Hmo =
amps × volts × pf × motor efficiency
746
Or, for three phase motors:
Hmo =
(3)0.5 × amps × volts × pf × motor efficiency
746
In both equations, amps and volts are the field test
measurement values and, in the case of three phase
motors, are the averages of the measured phase
values.
The fan power input is the motor power output minus
the power transmission loss, where applicable.
11.2.3 Calibrated motors. A calibrated motor may be
used to determine fan power input. When intending
to use this method, it is usually necessary to specify
in the motor purchase arrangements that the motor
be calibrated since an additional cost is normally
involved. Calibration data are similar to typical motor
performance data with the exception that, instead of
being merely typical, the calibration data represent
the performance of a specific motor, based on a test
of the motor. The motor is calibrated over a range of
operation. Electrical input data and other data
sufficient for the determination of power output are
obtained in the calibration. The calibration normally
provides data for operation at nameplate voltage, but
may include data for operation at voltages 10%
greater and 10% less than nameplate voltage. It is
important that the power supplied to the motor during
the field test be consistent with that used in its
calibration. The phase voltage should stable and
balanced, and the average should be within 2% of
the voltage at which the motor was calibrated. The
field test measurements and the calculations
involved in the determination of motor power output
are the same as those described in Section 11.2.2 for
use with typical motor performance data. The fan
power input is the motor power output minus the
power transmission loss, where applicable.
A calibrated motor provides accurate data to
determine motor power output. However, the cost of
the calibration is a limiting factor in the use of this
method in field tests. For low horsepower
applications, the fan manufacturer may be able to
calibrate a motor.
11.2.4. Torquemeters. Another method to determine
fan power input involves the use of a torquemeter
installed between the fan and the driver. The use of a
torquemeter requires some prearrangement with the
purchaser, who would normally have specified such
equipment, so that site conditions can be altered to
accommodate its installation. The torquemeter is
extremely limited in field test application. This is due
mainly to is high cost and the cost of its installation.
In addition, the length of the shut down time and the
revisions to site conditions required for its installation
are usually undesirable. For practical considerations,
it is not normally used in cases where the fan is belt
driven and where the fan impeller is installed directly
on the motor shaft.
11.3 Power measuring instruments
Measurement of current, voltage, watts, and power
factor can be obtained by using an industrial type
power analyzer of good quality. This type of
instrument is available with accuracies of 1% full
scale for volts, amps and power factor, and 2% full
scale for watts. Normally, the higher levels of
accuracy requirements can be met by using this type
of instrument, providing the measurements are well
up on the scales.
In many cases, accuracy level requirements will
permit the use of a clip-on type ammeter-voltmeter.
Clip-on instruments with accuracies of 3% full scale
are available.
11.4 Power transmission losses
Several types of power transmission equipment are
used in driving fans. Those in which power
transmission losses should be considered in the
determination of fan power input include belt drives,
gear boxes, fluid drives, and electromechanical
couplings.
Information as to whether the fan power input ratings
include power transmission losses is included in the
published performance ratings or is otherwise
available from the fan manufacturer. It is important
that this be established and that the fan power input
be determined accordingly in order to provide a valid
comparison of field test results to the fan
performance ratings. In most cases, fan power input
ratings do not include power transmission losses.
11.4.1 Estimating belt drive losses. In view of the
lack of published information available for use in
calculating belt drive losses, a graph is included in
Annex L for this purpose. As indicated in the graph,
belt drive loss, expressed as a percentage of motor
power output, decreases with increasing motor
power output and increases with increasing speed.
This graph is based on the results of over 400 drive
loss tests provided to AMCA by drive manufacturers.
The graph serves as a reasonable guide in
evaluating belt drive losses. The calculation of belt
drive loss, using this graph, is included in many of the
examples in Annex A.
Field Performance Measurement | 157
11.4.2 Estimating other transmission losses. For
other types of power transmission equipment, consult
the fan manufacturer to establish whether
transmission losses are included in the fan ratings,
and if so, request the magnitudes of the losses
allowed in the ratings. Otherwise, it will be necessary
to consult the manufacturer of the power
transmission equipment for the information regarding
transmission losses.
11.5 Accuracy
The uncertainty analyses presented in Annex T
indicate that the uncertainties in fan power input
determinations are expected to range from 4% to 8%.
This range is based on considerations of the
conditions encountered in most field test situations,
estimated accuracies of the various test methods
presented in this publication and allowances for
uncertainties in the determinations of power
transmission losses.
12. Fan Speed
12.1 Speed measuring instruments
Measure speed with a revolution counter and
chronometer, a stroboscopic tachometer, an
electronic counter-timer, or any other precision type
tachometer which has a demonstrated accuracy of
0.5% of the measured value. Friction driven and
magnetic type pickups should not be used in low fan
power ranges where they can influence speed and
fan power input measurements.
12.2 Speed measurements
Establish the speed by averaging a minimum of three
measurements made during the test determination
period. The variation in the measurements should not
exceed 1% for any single point of operation.
13. Densities
13.1 Locations of density determinations
Determine the densities of the gas stream for Plane
1, the fan inlet; and for Plane 3, the velocity pressure
measurement plane. In addition, the density at Plane
2, the fan outlet, must be determined whenever the
fan total pressure, the fan velocity pressure, or an
SEF at the outlet side of the fan is required.
13.2 Data required at each location
The pressure and temperature of the gas stream
must be obtained for each plane at which a density
158 | Field Performance Measurement
determination is required. The pressures at Planes 1
and 2 are based on the static pressure
measurements made for the purpose of determining
the fan static pressure. The pressure at Plane 3 is
obtained by averaging static pressure measurements
made concurrent with the velocity pressure
measurements made in a traverse of Plane 3. The
absolute pressure at a plane is calculated by using
the static pressure at the plane and the barometric
pressure. For this reason, it is important that the
barometric pressure be determined for the
atmosphere to which static pressure measurements
are referred. The temperatures used in density
determinations are measured at the planes of
interest.
13.3 Additional data
Additional data required in the determination of
density depends on the gas stream as indicated
below:
1) For air, the wet-bulb temperature is required
unless it is otherwise known that the air is
saturated with water vapor or that the water
vapor content of the air is insignificant. It should
be noted that incorrect assumptions as to
whether the air is dry or saturated can result in
substantial errors in density determinations.
2) For gases other than air, the normal procedure is
to rely on process personnel for the data
necessary to determine the density of the gas.
The information provided will include density or
data sufficient to calculate the density, which
should be for stated conditions of temperature
and pressure.
13.4 Density values
Gas stream density can be established when the
pressure, temperature, and additional data, as
indicated in Section 13.3, have been obtained.
Procedures for establishing density are described in
the examples in Annex M and are further illustrated in
the field test examples in Annex A.
Although the pressure and temperature of the gas
stream must be obtained for each plane at which a
density value is required, it is usually necessary to
obtain additional data, such as the wet-bulb
temperature, for only one plane in order to establish
the densities at all planes. The densities at the planes
for which the additional data is not obtained can be
calculated, providing the gas stream does not change
composition or undergo a change in phase between
planes. The calculation is based on density being
directly proportional to absolute pressure and
inversely proportional to absolute temperature.
13.4.1 Example calculation - ρ3 from ρ1. Use Figure
N.1 of Annex N to establish the density of air at Plane
1 based on the test determinations of barometric
pressure, pb, and the following Plane 1 values:
Ps1, static pressure, in. wg
td1, dry-bulb temperature, °F
tw1, wet-bulb temperature, °F
The following data are obtained for Plane 3:
Ps3, static pressure, in. wg
td3, dry-bulb temperature, °F
Calculate the density at Plane 3 as follows:
⎛ P + 13.6 pb ⎞ ⎛ t d1 + 460 ⎞
ρ3 = ρ1 ⎜ s3
⎟
⎟⎜
⎝ 13.6 p1 ⎠ ⎝ t d3 + 460 ⎠
thermometer should be accurate within 5°F of the
measured value and readable to 5°F or finer.
The temperature determination should be
representative of the average temperature of the gas
stream throughout the plane of interest. When the
temperature varies with time or temperature
stratification exists at the measurement plane,
several temperature measurements may be
necessary in order to obtain a representative
average. At elevated temperatures, the thermometer
may have to be shielded to prevent radiation effects
from exposed heat sources.
Locate the wet-bulb thermometer downstream from
the dry-bulb thermometer in order to prevent the drybulb temperature measurement from being adversely
affected. The wet-bulb thermometer wick should be
clean, closely fitted, and wetted with fresh water. The
velocity of the air over the wick should be between
700 and 2000 fpm. Use a sling psychrometer to
obtain dry and wet-bulb air temperature
measurements at the fan inlet for free inlet fans.
Where:
13.6 Barometric pressure
p1 = the absolute pressure, in. Hg at Plane 1,
calculated as follows:
p1 = pb + (Ps1/13.6)
In this manner, ρ3 can be calculated without having to
measure the wet-bulb temperature at Plane 3. These
equations can be used for gases other than air and
can be adapted for use in calculations involving any
two planes, subject to the limitations noted earlier.
In the example calculation of ρ3, pb is determined for
the atmosphere to which the measurements of Ps1
and Ps3 are referred. Refer static pressure
measurements to a common atmosphere. When
the pressures cannot be referred to a common
atmosphere, the absolute pressure for each plane is
calculated by using the static pressure measurement
at the plane and the barometric pressure for the
atmosphere to which the static pressure
measurement is referred. However, for the purposes
of accuracy, static pressure measurements that are
used in the determination of fan static pressure must
be referred to a common atmosphere.
13.5 Temperatures
Measure temperatures with mercury-in-glass, dial, or
thermocouple type thermometers. For temperatures
through 220°F, the thermometer should be accurate
within 2°F of the measured value and readable to 1°F
or finer. For temperatures above 220°F, the
Use a portable aneroid barometer for field test
determinations of barometric pressure when an
acceptable site barometer is not available. The
barometer should be accurate within 0.05 in. Hg of
the measured value. Determine the test value of
barometric pressure by averaging measurements
made at the beginning and end of the test period.
When the test value of barometric pressure is to be
based on data obtained from a nearby airport, it is
important that the data include the barometric
pressure for the airport site and the elevation for
which the pressure was determined (often the
barometric pressure is corrected to sea level). This
pressure value must then be corrected to the test site
elevation.
Barometric
pressure
decreases
approximately 0.1 in. Hg for every 100 ft increase in
elevation
13.7 Accuracy
As indicated in Annex T, uncertainties in density
determinations are expected to be less than 3%.
However, care must be exercised in obtaining
representative test measurements in order to prevent
the uncertainties from exceeding this value.
14. Conversion Calculations
Generally, the test fan will be operating at a speed
and inlet density that are somewhat different from the
Field Performance Measurement | 159
fan performance rating values of fan speed and inlet
density. In order to provide a common basis for
comparing the field test results to the fan
performance ratings, each of these two items must
be the same in both sets of data. This can be
accomplished by converting the results of the field
test to the speed and density conditions of the fan
performance ratings. The equations for the
conversion are as follows.
Qc = Q (Nc / N)
Psc = Ps (Nc / N)2 (ρc / ρ)
Ptc = Pt (Nc / N)2 (ρc / ρ)
Pvc = Pv (Nc / N)2 (ρc / ρ)
Hc = H (Nc / N)3 (ρc / ρ)
Where the subscript c designates values converted
to specified conditions, and items without the
subscript c are field test values.
These conversion equations do not account for the
effect of the compressibility of the gas stream.
However, since the test fan usually operates at
conditions of speed and inlet density that are
reasonably close to the quoted fan performance, the
conversion calculations usually result in small
changes from field test values and the effect of the
compressibility of the gas stream is considered to be
negligible. Where test conditions are considerably
different than design conditions, the effect of
compressibility may need to be considered.
Work
required
to
accommodate
test
measurements (drilling of traverse holes,
installation of static pressure taps and
thermometer wells, etc.) should be completed
prior to the test date.
4) System Effect Factors, if any, must be
established prior to the conduct of the test.
5) The expected test uncertainties must be agreed
upon prior to the test (see Annex T).
6) Responsibility for the cost of the test or any fansystem modifications required as a result of the
test should be established.
7) Prior to testing, an inspection must be made to
ensure that the fan is installed in accordance with
the fan manufacturer’s recommendations. The
duct system should also be inspected for
compliance with design specifications, conditions
of filters, abnormal duct restrictions, etc.
8) The majority of fan field performance tests cover
a single point of operation, namely, the design
duty. If it is deemed necessary to cover several
points of operation, provision must be made in
advance for changing the system resistance. The
means used to vary the system resistance must
not cause adverse flow conditions in the vicinities
of the fan and measurement planes. If the fan
cannot be tested at the quoted system design
point, then it is sufficient for the evaluation of fan
field performance to establish the proximity of the
field test point to any portion of the fan
performance rating curve within the limitations of
the uncertainty analysis (see Annex T).
15. Test Preparation
15.1 The following items should be agreed upon by
all interested parties prior to the start of a field
performance test:
1) AMCA Publication 200, Air Systems, AMCA
Publication 201, Fans and Systems, and AMCA
Publication 202, Troubleshooting, should be
reviewed and implemented before starting the
field test.
2) Personnel conducting field tests on fans must be
technically competent and fully conversant with
all four parts of the AMCA Fan Application
Manual. The person responsible for conducting
the test should be designated and agreed upon
by all parties.
3) The test instrumentation and locations of test
measurement planes should be established.
160 | Field Performance Measurement
9) It must be established that the system remains
constant for the duration of the test. Modulating
dampers should be set in a fixed position, no
process changes shall be undertaken, etc.
Variable inlet vane controls or inlet box dampers
must be set in the full open position for the
duration of the test, except when testing for
control characteristics.
10) All precautions to ensure the safety of test
personnel must be observed.
11) The fan-system should be operated for a length
of time sufficient to ensure steady state
conditions prior to the start of the test.
12) It is advisable that representatives of all parties
interested in the test results be present at the
time of the test to cover their areas of
responsibility.
15.2 It is recommended that as a minimum, the
following equipment be taken to or be otherwise
available at the job site:
1) Pitot-static tubes of suitable lengths for the
maximum duct size to be traversed.
Considerations should be given to the use of a
double reverse tube in dirty atmospheres.
2) Manometers suitable for measuring static
pressures. Manometer fluids other than water are
acceptable, provided the specific gravity is
known. A spare bottle of manometer fluid is
advisable.
3) Inclined manometer suitable for measuring
velocity pressures.
4) Flexible tubing of suitable length to enable
manometers to be installed at a convenient
location.
5) Tubing couplings and “T” type tubing connectors.
6) Thermometers to cover the range of anticipated
temperatures.
7) Sling psychrometer for obtaining dry-bulb and
wet-bulb temperatures.
8) Clip-on ammeter-voltmeter, power analyzer, or
other
suitable
electrical
measurement
instruments for the determination of fan power
input.
9) Fan speed measurement instrument.
10) Aneroid barometer.
11) Flashlight, tape, measuring rule, hand tools,
coveralls, etc.
12) Test data sheets, calculator, and necessary
drawings.
13) Complete AMCA Fan Application Manual
containing Publications 200, 201, 202, and 203.
2) Static and total pressure manometer tubing must
be “pinched off” prior to inserting or removing the
Pitot-static tube from the test duct. Release both
legs of the tubing simultaneously after the Pitotstatic tube is inside the test duct and properly
oriented. Failure to release simultaneously may
result in manometer fluid being blown from the
manometer.
3) Loop the manometer tubing well above the
manometer so that any fluid which is
inadvertently blown from the gauge will drain
back into the manometer.
4) The Pitot-static tube is intended for measuring
pressures in relatively clean gases. When using
Pitot-static tubes in dirty, wet, or corrosive
atmospheres, both legs of the Pitot-static tube
must be cleaned out frequently during the test.
Since fan pressure readings are never strictly
steady, absence of fluctuations is an
indication of a plugged Pitot-static tube.
Consider using a double reverse tube in these
situations.
5) When making measurements in wet gas
streams, continually check for the presence of
moisture in the tubing. Clear plastic tubing is
ideal from this standpoint. If moisture collects in
the tubing, immediately remove the Pitot-static
tube and clean the inside of the tubing and Pitotstatic tube before proceeding with the test.
6) Before performing any work inside a fan,
ductwork, or other system components, make
certain that the fan motor starter is “locked out.”
7) The area at the plane of flow measurement
should be measured internally to account for
internal insulation or other obstructions.
8) Do not rely on damper control indicators to
ensure that dampers are fully open. Check
visually.
9) Measure temperatures on both sides of double
inlet fans as temperature differences may exist
between each side.
16. Precautions
The following precautions should be observed when
conducting a field test:
1) Connect the Pitot-static tube to the manometers
according to anticipated pressures, i.e., whether
the pressures are positive or negative, and the
magnitudes of pressures.
10) When measuring in high temperature, corrosive
or explosive atmospheres, instruments should be
selected for suitability for such atmospheres.
17. Typical Fan-System Installations
A fan assembly may include any number of
appurtenances: variable inlet vanes, inlet boxes, inlet
box dampers, outlet dampers, inlet screens, belt
Field Performance Measurement | 161
guards, inlet bells, diffusers (evasés). Alternately,
these items may be included in the fan-system
installation, but not be a part of the fan assembly. In
order to determine the proper field test procedure
and to provide a valid basis for comparing field test
results to the fan performance ratings, it is important
to establish which of these items are considered a
part of the fan and which are considered a part of the
system. The fan performance ratings may be
assumed to include the appurtenances that are
established as being a part of the fan assembly.
The locations of the fan inlet and fan outlet depend
on whether specific appurtenances are considered
to be a part of the fan assembly. If the assembly
includes an inlet box, the fan inlet is the inlet to the
inlet box. For a fan assembly that includes a diffuser,
the fan outlet is the outlet of the diffuser.
In the case of heating, ventilating, and airconditioning equipment, the field test procedure will
depend on whether the equipment is a factory
assembled central station unit, a built-up unit, or a
packaged unit (see Section 17.4).
The performance ratings for a fan that includes inlet
box dampers, variable inlet vanes or outlet dampers
cover operation of the fan with these items in the full
open positions. In order to be able to compare the
field test results to the fan performance ratings, it is
essential that these items be fixed in their full open
positions for the duration of the test. In addition, when
the loss through a damper must be calculated, it is
essential that the damper blades be fixed in their full
open positions during the test since this is the
condition on which the damper pressure loss ratings
are based. This consideration arises when a damper,
which is not considered a part of the fan is located
between a static pressure measurement plane and
the fan. In order to determine the fan static pressure,
the loss through the damper must be calculated. In
these cases, the calculation of the loss is based on
the performance ratings for the damper.
17.1 Free inlet, free outlet fans
It is difficult to achieve an accurate field test of a free
inlet, free outlet fan. The most obvious problem is the
lack of a suitable location for the velocity pressure
measurement plane. In addition, in the case of
ventilators that supply or exhaust air from a buildingthe most commonly encountered applications of free
inlet, free outlet fans-it is extremely difficult to define,
set, and maintain for the duration of the test the
“normal” system condition. Items affecting the system
include:
162 | Field Performance Measurement
a) The operations of ovens, furnaces, paint booths,
air conditioning equipment, other fans, and
similar items that may supply or exhaust air from
the building in intermittent or modulating
fashions.
b) The use of doors providing access to the
building. The effect is most significant when large
doors that are normally closed are kept open for
extended periods such as in loading operations.
c) The velocity and direction of the wind outside the
building, particularly in conjunction with the item
immediately above and as it may affect the flow
of air from the outlet of the ventilator.
d) The use of interior doors that my restrict the flow
of air from areas normally expected to be
ventilated.
Assuming that these difficulties can be resolved and
the desired system is fixed for the duration of the test,
determine the fan performance by using one of the
following methods:
1) Make field test measurements sufficient for
determining fan static pressure, fan power input,
fan speed, and the density of the air at the fan
inlet. In this method for testing a free inlet, free
outlet fan, the fan static pressure is calculated as
the static pressure on the outlet side of the fan
less the static pressure on the inlet side of the
fan: Ps = Ps2 - Ps1. The static pressure
measurements involved must be referred to the
same atmospheric pressure and made at
locations sufficiently distant from the fan inlet and
outlet so as to be unaffected by the velocity of the
air entering and leaving the fan. Using the fan
manufacturer’s certified performance ratings,
draw a performance curve for the fan for
operation at the test values of fan speed and
entering air density. Determine the fan air flow
rate by entering this curve at the test values of
fan static pressure and fan power input (see
Example 5C in Annex A).
2) Use the method as described above with the
exception that the performance curve is
established by a laboratory test of the fan,
conducted in accordance with AMCA Standard
210. For the laboratory test, the fan must be set
up in a manner that duplicates the field
installation conditions. That is, all appurtenances
must be in place and any restrictions or
obstructions to the free flow of air into the fan
inlet and away from the fan outlet must be
accurately duplicated in the laboratory test setup.
3) Install a duct on the inlet side of the fan for the
purpose of providing a location for the velocity
pressure measurement plane. All of the test
measurements and calculations in this method
for testing a free inlet, free outlet fan are the
same as those required for a fan with a ducted
inlet and a free outlet. The cross-sectional shape
and area of the duct, which is temporarily
installed for purposes of the test, should be
selected on the basis of minimizing its
interference with the flow of air into the fan inlet
while providing velocity pressure of magnitudes
that can be accurately measured. The length of
the duct should be a minimum of twice its
diameter or equivalent diameter, and the
entrance to the duct should be flared in order to
reduce the entrance loss. The velocity pressure
measurement plane should be located a
minimum of 1.5 diameters or equivalent
diameters downstream from the duct inlet. The
effect of this duct on the system is negligible, but
in changing the pattern of the flow of air into the
fan inlet, it may affect the performance of the fan
slightly. Applications of this method of test are
presented in Examples 5A and 5B in Annex A.
The equation for calculating fan static pressure
for this configuration is:
Ps = Ps2 - (Ps1 + Pv1)
17.2 Free inlet, ducted outlet fans
In the calculation of fan static pressure for this type of
fan-system configuration, the sum of the static
pressure at the fan inlet, Ps1, and the velocity
pressure at the fan inlet, Pv1, is considered to be
equal to the sum of the static pressure, Psx, and the
velocity pressure, Pvx, at a point sufficiently distant
from the fan inlet as to be in still air. At this point, the
static pressure is zero, and the velocity pressure in
still air is zero.
Ps1 + Pv1 = Psx + Pvx = 0
This consideration, which is the same as that used in
the methods for testing fans for performance rating
purposes, charges to the fan the losses incurred in
accelerating the air into the fan inlet and eliminates
inaccuracies which may occur in any attempt to
measure velocity pressure and static pressure at the
fan inlet. Since Ps1 + Pv1 = 0, the equation for
calculating fan static pressure for this configuration
is:
Ps = Ps2 + SEF 1 +SEF 2 + ... + SEF n
17.3 Ducted inlet, ducted outlet fans
In this type of fan-system configuration, there is no
special consideration in the calculation of fan static
pressure. The equation for this calculation is:
Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2 + ... + SEF n
In this configuration, the flow conditions on the inlet
side of the fan are usually more favorable for the
location of the velocity pressure measurement plane.
17.4 Ducted inlet, free outlet fans
In this type of fan-system configuration, the static
pressure at the fan outlet, Ps2, is zero gauge
pressure, referred to the atmospheric pressure in the
region of the fan outlet. However, the gas stream may
be discharging from the fan into a region in which the
atmospheric pressure is somewhat different from that
to which all other pressure measurements are
referred. When this possibility exists, it is essential
that the static pressure measurements in the region
of the fan outlet be referred to the same atmospheric
pressure as used in all other pressure
measurements.
Ps = -Ps1 - Pv1 + SEF 1 + SEF 2 + ... + SEF n
17.5 Air handling units
This category consists of draw-through and blowthrough types of equipment assemblies used in
heating, ventilating, and air-conditioning applications.
In addition to fans, these equipment assemblies may
include any number of combinations of coils, filters,
access sections, humidifiers, mixing boxes, dampers,
etc. Air handling units include packaged units, factory
assembled central station units, and built-up units.
The basis used in establishing the air performance
ratings for each of these unit types is described
below. It is important that the field test method
correspond to the rating method in each case.
17.5.1 Packaged units. This type of unit is supplied
and rated by the manufacturer as an assembly. The
static pressures at the inlet and outlet to the
assembly and the velocity pressure at the inlet to the
assembly are used in calculating the static pressure
generated by this type of air handling unit. See
Examples 4C and 4D in Annex A.
17.5.2 Factory assembled central station units.
The air performance ratings for this type of unit are
based on the operation of the fan section assembly
only, but include the effects of the air flow conditions
entering and leaving the fan section which are
created by accessory equipment such as plenums,
Field Performance Measurement | 163
coils, filters, mixing boxes, etc. The fan section
assembly includes the fan and the cabinet in which
the fan has been installed. The accessory items are
considered to be included in the system in which the
fan section operates. The static pressure and the
velocity pressure at the inlet of the fan section and
the static pressure at the fan section outlet, which
coincides with the fan outlet, are used in calculating
the static pressure generated by the fan section
assembly. See examples 4B and 4E in Annex A.
17.5.3 Built-up units. Built-up units are similar to
factory assembled central station units, except that in
built-up units, the components are normally obtained
from a number of equipment suppliers and the unit is
assembled at the installation site. The fans which are
used in built-up units are rated as free-standing,
unencumbered by the cabinets in which they are
installed. In the field test determination of the
performance of the fan, the static pressure and
velocity pressure at the fan inlet and the static
pressure at the fan outlet are used in calculating the
fan static pressure. An SEF that accounts for the
effect of the cabinet is normally included in this
calculation, and it may be necessary to include an
SEF to account for the conditions at the fan outlet.
See Example 4A in Annex A.
164 | Field Performance Measurement
Annex A. Field Test Examples
This annex contains examples of field tests. The examples are presented in detail and cover several types of fansystem combinations. Field test procedures are illustrated in a variety of situations. Portions of the procedures are
typical for all fan-system installations. Other portions of the procedures demonstrate methods for dealing with the
more difficult features encountered in some installations. Not all of the possible fan-system combinations are
included in the examples, but it is expected that the examples will provide sufficient guidance for dealing with those
cases not covered.
EXAMPLES OF FANS, INSTALLATION TYPE B: FREE INLET, DUCTED OUTLET
1A:
1B:
1C:
1D:
Centrifugal Forced Draft Fan
Centrifugal Forced Draft Fan with Inlet Silencers
Axial Forced Draft Fan with Inlet Silencers
Centrifugal Fans in Parallel
EXAMPLE OF FANS, INSTALLATION TYPE D: DUCTED INLET, DUCTED OUTLET
2A:
2B:
2C:
2D:
2E:
2F:
2G:
Utility Fan in a Ventilating System
Centrifugal Fan in a Sawdust Conveying System
Axial Fan in a Dryer System
Centrifugal Fan in a Scrubber System
Centrifugal Fan in a Process System
Axial Fan in a Ventilation System
High Pressure Centrifugal Fans in Series
EXAMPLES OF FANS, INSTALLATION TYPE C: DUCTED INLET, FREE OUTLET
3A:
3B:
3C:
3D:
Centrifugal Fan in an Exhaust System
Axial Fan in an Exhaust System
Centrifugal Fan in a Scrubber System
Centrifugal Roof Ventilator with Ducted Inlet
EXAMPLES OF AIR HANDLING UNITS
4A:
4B:
4C:
4D:
4E:
Centrifugal Fan in a Built-up Air conditioning Unit
Central Station Air Conditioning Unit, Factory Assembled Draw-Through Type
Packaged Air Conditioning Unit
Packaged Air Conditioning Unit
Central Station Air Conditioning Unit, Factory Assembled Blow-Through Type
EXAMPLES OF FANS, INSTALLATION TYPE A: FREE INLET, FREE OUTLET
5A:
5B:
5C:
Free Inlet, Free Outlet Roof Ventilator with temporary duct
Free Inlet, Free Outlet Propeller Fan with temporary duct
Free Inlet, Free Outlet Roof Ventilator as installed
Field Performance Measurement | 165
EXAMPLE 1A: CENTRIFUGAL FORCED DRAFT FAN
SEF 1
DIFFUSER
3
2
L
A2
VARIABLE
INLET VANES
SIDE VIEW
A3
OUTLET SIDE VIEW
LOCATIONS OF
PLANES 2 AND 3
ORIENTATION
OF PITOT TUBE
COMMENTS
1. The variable inlet vanes are considered part of the
fan. Performance ratings for fans with inlet vanes
cover operation with the inlet vanes in their full open
position. In order to be able to compare the test
results to the fan performance ratings, it is essential
that the inlet vanes be fixed in their full open positions
for the duration of the test.
2. Determine Pv3 by using the root mean square of
the velocity pressure measurements made in a
traverse of Plane 3, located near the end of the fan
diffuser (evasé). Determine Ps3 by averaging the
static pressure measurements made in the same
traverse. Procedures for the traverse are described in
Section 9.4. These velocity pressure and static
pressure measurements are susceptible to error due
to the turbulence existing in the region of the fan
outlet. In addition, it is undesirable to have Plane 3
located in a diverging airway. However, no other
more suitable location for Plane 3 exists in this
example. It is recommended that the Pitot-static tube
be oriented so that its nose is aligned with the
anticipated flow streams, particularly near the walls of
the diffuser, as shown in the diagram. Determine the
area of the traverse plane, A3, which is located at the
tip of the Pitot-static tube, as shown in the diagram,
not at the location of the Pitot-static tube access
holes in the diffuser.
166 | Field Performance Measurement
3. Measure td1 and tw1 in the path of the air flowing into
the fan inlets. Determine pb for the general vicinity of
the fan. Measure td3 in Plane 3. All of these
measurements are used in the determination of
densities at the various planes of interest.
4. Measure the fan speed and the motor amps, volts,
and, if possible, watts. Record all pertinent motor
nameplate data, including volts (NPV) and full load
amps (FLA). If the motor power output is to be
estimated by using the phase current method
described in Annex K, it is not necessary to measure
motor watts; however, it may be necessary to
disconnect the drive and measure the no load amps
(NLA) if the motor is not operating at or near its full
load point (refer to Annex K).
5. SEF 1 is due to the effect of insufficient length of
duct at the fan outlet. In order to calculate the value
of SEF 1, it is necessary to measure the length of the
outlet duct, L; the outlet area of the fan, A2; and the
blast area of the fan.
6. The sum of the static pressure, Ps1, and velocity
pressure, Pv1, at the inlets of a fan with unrestricted
inlets is considered to be equal to the sum of the
static pressure, Psx, and the velocity pressure, Pvx, at
a point sufficiently distant from the fan inlets as to be
in still air. At this point, the static pressure is zero, and
the velocity pressure in still air is zero.
GENERAL
Ps1 + Pv1 = Psx + Pvx = 0
VIVs in full open positions.
Fan direct connected to motor.
This consideration, which is the same as that used in
the methods for testing fans for performance rating
purposes, charges to the fan losses incurred in
accelerating the air into the fan inlets and eliminates
the inaccuracies which arise in any attempt to
measure the velocity pressure and static pressure at
the fan inlets. To calculate the fan static pressure:
DENSITIES
Ps = Ps2 - Ps1 - Pv1 + SEF 1
= Ps2 - (Ps1 + Pv1) + SEF1
td1 = 85°F
tw1 = 63°F
p1 = pb
= 28.91 in. Hg
CALCULATIONS
For fan inlet conditions of:
Since:
Ps1 + Pv1 = 0
Ps = Ps2 + SEF 1
Use Figure N.1 in Annex N to obtain ρ1 = 0.0701
lbm/ft3
The density at Plane 3:
7. In order to compare the test results to the quoted
fan curve drawn for operation at 1780 rpm and
0.0701 lbm/ft3 density, it is necessary to convert the
results to the specified conditions. In this case, the
test conditions are identical to the specified
conditions and no calculations are required.
OBSERVATIONS
⎛ P + 13.6 pb ⎞ ⎛ t d1 + 460 ⎞
ρ3 = ρ1 ⎜ s3
⎟
⎟⎜
⎝ 13.6 p1 ⎠ ⎝ t d3 + 460 ⎠
⎛ 14.4 + 13.6 × 28.91 ⎞ ⎛ 545 ⎞
= 0.0701⎜
13.6 × 28.91 ⎟⎠ ⎜⎝ 556 ⎟⎠
⎝
= 0.0712 lbm/ft 3
SITE MEASUREMENTS
In this case, ρ2 is considered to be equal to ρ3.
pb = 28.91 in. Hg
td1 = 85°F
tw1 = 63°F
td3 = 96°F
Ps3 = 14.4 in. wg
Pv3 = 1.52 in. wg
N = 1780 rpm
A2 = 11.94 ft2
A3 = 11.3 ft2
Blast Area = 7.76 ft2
L = 3 ft.
FLOW RATES
MEASURED MOTOR DATA
Volts = 570, 560, 572
= 567 av.
Amps = 160, 166, 163
= 163 av.
MOTOR NAMEPLATE DATA
200 hp, 3 phase, 60 hertz
575 volts, 1800 rpm, 181 FLA
V3 = 1096 (Pv3/ρ3)0.5
= 1096 (1.52/0.0712)0.5
= 5064 fpm
Q3 = V3A3
= 5064 × 11.3
= 57223 cfm
Q = Q1
= Q3 (ρ3/ρ1)
= 57223 (0.0712/0.0701)
= 58121 cfm
FAN POWER INPUT
Measured amps/FLA = (163/181)
= 0.90
= 90%
Annex K indicates that Equation A will provide a
reasonably accurate estimate of motor power output
for a 200 hp motor operating at 90% FLA.
Field Performance Measurement | 167
Hmo = 200 (163/181) (567/575)
= 178 hp
Since the fan is direct connected to the motor:
H
= Hmo
= 178 hp
SYSTEM EFFECT FACTOR
AMCA Publication 201-90, Figures 7.1 and 8.3
indicate the following calculations:
Q2 = Q3 (ρ3/ρ2)
= 57223 (0.0712/0.0712)
= 57223 cfm
V2 = (Q2/A2)
= (57223/11.94)
= 4793 fpm
Duct diameter equivalent to the diffuser outlet area:
De2 = 4 A2 / π
=
( 4 × 11.94 ) / π
= 3.9 ft.
Figure 8.3 shows that for velocities over 2500 fpm,
100% effective duct length is one duct diameter per
1000 fpm,
= De2 (V2/1000)
= 3.9 (4793/1000)
= 18.7 ft
L in % effective duct length
= (L/18.7) 100
= (3/18.7) 100
= 16%
Blast area ratio = Blast Area/A2
= 7.76/11.94
= 0.65
For blast area ratio of 0.65, and 16% effective duct
length, Figure 8.3 shows System Effect Curve U
applies. For 4793 fpm velocity and curve U, Figure
7.1 shows SEF 1 = 0.6 in. wg at 0.075 lbm/ft3. At
0.0712 lbm/ft3.
SEF 1 = 0.6 (0.0712/0.075)
= 0.57 in. wg
168 | Field Performance Measurement
FAN STATIC PRESSURE
Since A2 is greater than A3, there may be some
conversion of velocity pressure to static pressure
between Planes 3 and 2. However, the amount of
conversion will be very small relative to the static
pressure measured at Plane 3 and ignoring any
change in static pressure from Plane 3 to Plane 2 will
have no appreciable effect on the test results.
Therefore, Ps2 is considered equal to Ps3.
Ps = Ps2 + SEF 1
= 14.4 + 0.57
= 14.97 in. wg
CONVERSION TO SPECIFIED CONDITIONS
Qc = Q
= 58121 cfm
Psc = Ps
= 14.97 in. wg
Hc = H
= 178 hp
EXAMPLE 1B: CENTRIFUGAL FORCED DRAFT FAN WITH INLET SILENCERS
TEMPORARY
DUCT
DIFFUSER
STATIC
PRESSURE TAPS
3a
0.5 De
3b
SILENCERS
1
A2
SEF 1
VARIABLE INLET VANES
SIDE VIEW
OUTLET SIDE VIEW
2
COMMENTS
1. This fan, as supplied and rated by the
manufacturer, includes the variable inlet vanes and
inlet boxes, but does not include the silencers.
Performance ratings for fans with inlet vanes cover
operation with the inlet vanes in the full open
positions. In order to be able to compare the test
results to the fan performance ratings, it is essential
that the inlet vanes be fixed in their full open positions
for the duration of the test.
2. Determine Pv3a and Pv3b by using the root mean
square of the velocity pressure measurements made
in traverses of Planes 3a and 3b. A3a and A3b are the
areas traversed. Determine Ps3a and Ps3b by
averaging each of the two sets of static pressure
measurements made in the same traverses.
Procedures for traverses are described in Section
9.4. Ps3a and Ps3b are used in determining the density
at the traverse plane. A location for Plane 3
measurements may be obtained by installing ducts
on each silencer inlet, as shown in the diagram. The
ducts should be a minimum of one equivalent
diameter in length, and have flared inlets to reduce
entrance losses and provide more uniform velocity
profiles at the pressure measurement planes.
3. Measure Ps1a and Ps1b at locations close to the
entrances to the inlet boxes and in planes which are
substantially equal in area to the planes of the
entrances to the inlet boxes (Plane 1). Determine Ps2
by averaging the pressure measurements at each of
four static pressure taps located near the end of the
fan diffuser (evasé). See Annex E for details of static
pressure taps.
4. Measure td3 and tw3 near the inlet ducts. Determine
pb for the general vicinity of the fan. Measure td2 in
Plane 2. All of these measurements are used in the
determination of densities at the various planes of
interest.
5. Measure the fan speed and the motor amps, volts,
and if possible, watts. Record all pertinent motor
nameplate data, including volts (NPV) and amps
(FLA). If the motor power output is to be estimated by
using the phase current method described in Annex
K, it is not necessary to measure motor watts;
however, it may be necessary to disconnect the drive
and measure the no load amps (NLA) if the motor is
not operating at or near its full load point. Refer to
Annex K.
6. SEF 1 is due to the effect of there being no duct
at the fan outlet. In order to calculate the value of
SEF 1, it is necessary to measure the fan outlet area,
A2, and the blast area of the fan.
7. To calculate the fan static pressure:
Field Performance Measurement | 169
Ps = Ps2 - Ps1 - Pv1 + SEF 1
CALCULATIONS
Where:
DENSITIES
Pv1 = (Q/1096A1)2 ρ1
For Plane 3 conditions of:
8. In order to compare the test results to the quoted
fan curve drawn for operation at 1180 rpm and 0.075
lbm/ft3 density, it is necessary to convert the results
to the specified conditions. The basis for the
calculations is described in Section 14.
td3 = 85°F
tw3 = 58°F
OBSERVATIONS
SITE MEASUREMENTS
pb = 29.31 in. Hg
td2 = 93°F
td3 = 85°F
tw3 = 58°F
Ps1a = -1.20 in. wg
Ps1b = -1.30 in. wg
Ps2 = 10.1 in. wg
Ps3a = -0.65 in. wg
Ps3b = -0.70 in. wg
Pv3a = 0.61 in. wg
Pv3b = 0.62 in. wg
N = 1180 rpm
A1a = A1b
= 12.5 ft2
A2 = 18 ft2
A3a = A3b
= 12.5 ft2
Blast Area = 13.5 ft2
MEASURED MOTOR DATA
Volts = 460, 455, 465
= 460 av
Amps = 257, 256, 258
= 257 av
MOTOR NAMEPLATE DATA
200 HP, 3 phase 60 hertz
460 volts, 1180 rpm, 285 FLA
GENERAL
VIVs in full open positions. Fan direct connected to
motor. The motor manufacturer advises that this
motor type has a peak efficiency of 91% at a power
factor of approximately 0.89.
Ps3 = (Ps3a + Ps3b)/2
= (-0.65 - 0.70)/2
= -0.675 in. wg
p3 = pb + (Ps3/13.6)
= 29.31 + (-0.675/13.6)
= 29.26 in. Hg
Use Figure N.1 in Annex N to obtain ρ3 = 0.0712
lbm/ft3
It is assumed that the temperature at Plane 1 are the
same as those at Plane 3. The density at Plane 1:
⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞
ρ1 = ρ3 ⎜ s1
⎟⎜
⎟
⎝ 13.6 p3 ⎠ ⎝ t d1 + 460 ⎠
⎛ −1.25 + 13.6 × 29.31 ⎞ ⎛ 545 ⎞
= 0.0712 ⎜
⎟ ⎜ 545 ⎟
13.6 × 29.26
⎝
⎠⎝
⎠
= 0.0711 lbm/ft 3
The density at Plane 2:
⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞
ρ2 = ρ3 ⎜ s2
⎟⎜
⎟
⎝ 13.6 p3 ⎠ ⎝ t d2 + 460 ⎠
⎛ 10.1 + 13.6 × 29.31 ⎞ ⎛ 545 ⎞
= 0.0712 ⎜
⎟⎜
⎟
⎝ 13.6 × 29.26 ⎠ ⎝ 553 ⎠
= 0.0721 lbm/ft 3
FLOW RATES
V3a = 1096 (Pv3a/ρ3)0.5
= 1096 (0.61/0.0712)0.5
= 3208 fpm
Q3a = V3aA3a
= 3208 × 12.5
= 40100 cfm
V3b = 1096 (Pv3b/ρ3)0.5
= 1096 (0.62/0.0712)0.5
= 3234 cfm
Q3b = V3bA3b
= 3234 × 12.5
= 40425 cfm
170 | Field Performance Measurement
Q3 = Q3a + Q3b
= 40100 + 40425
= 80525 cfm
Q = Q1
= Q3 (ρ3/ρ1)
= 80525 (0.0712/0.0711)
= 80638 cfm
SYSTEM EFFECT FACTOR
AMCA Publication 201-90, Figures 7.1 and 8.3
indicate the following calculations:
Q3 (ρ3/ρ2) = 80525 (0.0712/0.0721)
= 79520 cfm
(Q2/A2)
FAN POWER INPUT
= (79520/18)
= 4418 fpm
Measured amps/FLA = (257/285)
= 0.90
= 90%
Blast area ratio = Blast Area/A2
= 13.5/18
= 0.75
Annex K indicates that Equation A will provide a
reasonably accurate estimate of motor power output
for a 250 hp motor operating at 90% FLA.
For a blast area ratio of 0.75, and no duct, Figure 8.3
shows System Effect Curve T applies. For 4418 fpm
velocity and curve T, Figure 7.1 shows SEF 1 = 0.65
in. wg at 0.075 lbm/ft3. At 0.0720 lbm/ft3:
Hmo = 250 (257/285) (460/460)
= 225 hp
As a check of this value, using the motor efficiency
data and the appropriate equation in Section 11.2.2:
3 × 257 × 460 × 0.89 × 0.91
Hmo =
746
= 222 hp
Since the motor is not fully loaded, the power factor
and efficiency may be less, which would reduce Hmo
as calculated using the second method. However,
this is a reasonable check. The value of Hmo is
selected to be the average of the two results:
SEF 1 = 0.65 (0.0721/0.075)
= 0.62 in. wg
FAN STATIC PRESSURE
Pv1 = (Q1/1096 A1)2
= (80638/1096 × 25)2 0.0711
= 0.62 in. wg
Ps = Ps2 - Ps1 - Pv1 + SEF 1
= 10.1 - (-1.25) - 0.62 + 0.62
= 11.33 in. wg
CONVERSION TO SPECIFIED CONDITIONS
Hmo = 224 hp
Qc = Q
= 80638 cfm
Since the fan is direct-connected to the motor, there
is no drive loss, and:
Psc = 11.33 (0.075/0.0711)
= 11.95 in. wg
H = Hmo
= 224 hp
Hc = 224 (0.075/0.0711)
= 236 hp
Field Performance Measurement | 171
EXAMPLE 1C: AXIAL FORCED DRAFT FAN WITH INLET SILENCER
PLANE 3
LOCATION
3
TEMPORARY
SHORT DUCT
STATIC PRESSURE TAPS
SILENCER
TRANSITION
0.5 De
5
INLET
BOX
1
DIFFUSER
SECTION
2
INNER
CYLINDER
L
SIDE VIEW
GUIDE VANES
COMMENTS
1. This is a variable pitch axial flow fan. The fan
assembly, as supplied and rated by the manufacturer,
includes the inlet box and diffuser section, but does
not include the silencer. It is essential that the blade
pitch angle be fixed for the duration of the test. This
blade angle should be agreed upon by all interested
parties.
2. A temporary short duct is installed upstream of the
silencer to establish Plane 3 in which more uniform
pressures can be obtained. The duct should be a
minimum of one equivalent diameter in length, and
have a flared inlet to reduce entrance losses and
provide a more uniform velocity profile at the
pressure measurement plane. Determine Pv3 by
using the root mean square of the velocity pressure
measurements made in a traverse of Plane 3. Ps3 is
determined by averaging the static pressure
measurements made in the same traverse.
Procedures for traverses are described in Section
9.4. Ps3 is used in determining the density at the
traverse plane.
3. Measure Ps1 at a location close to the entrance to
the inlet box and in a plane which is substantially
equal in area to the plane of the entrance to the inlet
box (Plane 1). Determine Ps5 by averaging the
pressure measurements at each of four static
pressure taps located near the end of the fan diffuser.
172 | Field Performance Measurement
See Annex E for details of static pressure taps. In this
example, Ps2 is considered to be equal to Ps5.
4. Measure td3 and tw3 near the entrance to the short
inlet duct. Determine pb for the general vicinity of the
fan. Measure td5 in Plane 5. All of these
measurements are used in the determination of
densities at the various planes of interest.
5. Measure the fan speed and the motor amps, volts,
and if possible, watts. Record all pertinent motor
nameplate data, including volts (NPV) and full load
amps (FLA). If the motor power output is to be
estimated by using the phase current method
described in Annex K, it is not necessary to measure
motor watts; however, it may be necessary to
disconnect the drive and measure the no load amps
(NLA) if the motor is not operating at or near its full
load point. Motor performance data, supplied by the
motor manufacturer, are used in the determination of
motor power output for this example.
6. SEF 1 is due to the effect of insufficient length of
duct between the diffuser outlet and the elbow
downstream of the diffuser. In order to calculate the
value of SEF 1, it is necessary to measure the length
of the transition, L, and the outlet area of the diffuser,
A2.
7. To calculate the Fan Static Pressure:
CALCULATIONS
Ps = Ps2 - Ps1 - Pv1 + SEF 1
DENSITIES
Where:
For Plane 3 conditions of:
Pv1 = Pv3 (A3/A1)2 (ρ3/ρ1)
td3 = 68°F
tw3 = 62°F
8. Axial fans are often rated in Fan Total Pressure.
Computation of Fan Total Pressure is illustrated in the
CALCULATIONS section of this example.
9. In order to compare the test results to the quoted
fan curve drawn for operation at 880 rpm and 0.0740
lbm/ft3 density, it is necessary to convert the results
to the specified conditions. In this case, the test
conditions are identical to the specified conditions
and no calculations are required.
OBSERVATIONS
SITE MEASUREMENTS
pb = 29.8 in. Hg
td3 = 68°F
tw3 = 62°F
td5 = 88°F
Ps1 = -1.80 in. wg
Ps3 = -1.40 in. wg
Ps5 = 20.8 in. wg
Pv3 = 1.30 in. wg
N = 880 rpm
A1 = 170.3 ft2
A2 = 176 ft2
A3 = 170.3 ft2
A5 = A2
L = 15 ft
MEASURED MOTOR DATA
Volts = 4000, 4000, 4100
= 4033 av
Amps = 450, 445, 448
= 448 av
MOTOR NAMEPLATE DATA
4000 hp, 3 phase 60 hertz
4000 volts, 900 rpm, 520 FLA
GENERAL
p3 = pb + (Ps3/13.6)
= 29.8 + (-1.40/13.6)
= 29.70 in. Hg
Use Figure 20 in Annex N to obtain ρ3 = 0.0744
lbm/ft3
It is assumed that td1 = td3. The density at Plane 1:
⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞
ρ1 = ρ3 ⎜ s1
⎟⎜
⎟
⎝ 13.6 p3 ⎠ ⎝ t d1 + 460 ⎠
⎛ −1.8 + 13.6 × 29.8 ⎞ ⎛ 528 ⎞
= 0.0744 ⎜
⎟⎜
⎟
⎝ 13.6 × 29.70 ⎠ ⎝ 528 ⎠
= 0.0743 lbm/ft 3
The density at Plane 2:
ρ 2 = ρ5
⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞
= ρ3 ⎜ s5
⎟⎜
⎟
⎝ 13.6 p3 ⎠ ⎝ t d5 + 460 ⎠
⎛ 20.8 + 13.6 × 29.8 ⎞ ⎛ 528 ⎞
= 0.0744 ⎜
⎟⎜
⎟
⎝ 13.6 × 29.70 ⎠ ⎝ 548 ⎠
= 0.0756 lbm/ft 3
FLOW RATE
V3 = 1096 (Pv3/ρ3)0.5
= 1096 (1.3/0.0744)0.5
= 4581 fpm
Q3 = V3A3
= 4581 × 170.3
= 780144 cfm
Q = Q1
= Q3 (ρ3/ρ1)
= 780144 (0.0744/0.0743)
= 781194 cfm
Fan direct connected to motor. Motor performance
data at operating load, as supplied by motor
manufacturer: 0.88 power factor, 95% efficiency.
Field Performance Measurement | 173
FAN POWER INPUT
Hmo =
3 × volts × amps × power factor × efficiency
746
3 × 4033 × 448 × 0.88 × 0.95
746
= 3507 hp
=
SEF 1 = 0.32 (0.0756/0.075)
= 0.32 in. wg
FAN STATIC PRESSURE
Pv1 = Pv3 (A3/A1)2 (ρ3/ρ1)
= 1.3 (170.3/170.3)2 (0.0744/0.0743)
= 1.30 in. wg
Since the fan is direct connected to the motor, there
is no drive loss, and:
Ps2 = Ps5
= 20.8 in. wg
H = Hmo
= 3507 hp
Ps = Ps2 - Ps1 - Pv1 + SEF 1
= 20.8 - (-1.80) - 1.30 + 0.32
= 21.62 in. wg
SYSTEM EFFECT FACTOR
FAN TOTAL PRESSURE
AMCA Publication 201-90, Figures 7.1, 8.1, and 8.4
indicate the following calculations:
Q2 = Q3 (ρ3/ρ2)
= 780144 (0.0744/0.0756)
= 767761 cfm
V2 = (Q2/A2)
= (767761/176)
= 4362
Duct diameter equivalent to the diffuser outlet area:
De2 = 4 A2 / π
=
( 4 × 176 ) / π
= 15 ft.
Figure 8.1 shows that for velocities over 2500 fpm,
100% effective duct length is one duct diameter for
every 1000 fpm:
= De2 (V2/1000)
= 15 (4362/1000)
= 65.43 ft.
L in % effective duct length
= (L/65.43) 100
= (15/65.43) 100
= 23%
For 23% effective duct length and a vaneaxial fan
with a 2 piece elbow, Figure 8.4 shows System Effect
Curve V applies. For 4362 fpm velocity and curve V,
Figure 7.1 shows SEF 1 = 0.32 in. wg at 0.075
lbm/ft3. At 0.0756 lbm/ft3.
174 | Field Performance Measurement
Pt1 = Ps1 +Pv1
= -1.8 + 1.30
= -0.50 in. wg
Pv2 = Pv3 (A3/A2)2 (ρ3/ρ2)
= 1.3 (170.3/176)2 (0.0744/0.0756)
= 1.20 in. wg
Pt2 = Ps2 + Pv2
= 20.8 + 1.20
= 22.00 in. wg
Pt = Pt2 - Pt1 + SEF 1
= 22.00 - (-0.50) + 0.32
= 22.82 in. wg
Also:
Pt = Ps + Pv
Pv = Pv2
= 1.20 in. wg
Pt = 21.62 + 1.20
= 22.82 in. wg
CONVERSION TO SPECIFIED CONDITIONS
Qc = Q
= 781194 cfm
Psc = Ps
= 21.62 in. wg
Ptc = Pt
= 22.82 in. wg
Hc = H
= 3507 hp
EXAMPLE 1D: CENTRIFUGAL FANS IN PARALLEL
3
STATIC PRESSURE TAPS
OUTLET DAMPER
2
SEF 1
PLENUM
1
PLAN VIEW
1
SIDE VIEW
COMMENTS
1. Each of the fans, as supplied and rated by the
manufacturer, includes an outlet damper.
Performance ratings for fans with outlet dampers
cover operation with the outlet damper in the full
open position. In order to be able to compare the test
results to the fan performance ratings it is essential
that the outlet dampers be fixed in the full open
positions for the duration of the test.
2. In this example, there are no suitable locations for
traverse planes for use in determining directly the
flow rate for each fan. The alternative is to determine
the total flow rate and since the fans and their
operating speeds are alike, assume that each fan
delivers a flow rate proportional to its actual speed.
Determine Pv3 by using the root mean square of the
velocity pressure measurements made in a traverse
of Plane 3, located near the end of a straight run of
duct, such as shown in the diagram. Determine Ps3 by
averaging the static pressure measurements made in
the same traverse. Procedures for traverses are
described in Section 9.4. Ps3 is used in determining
the density at the traverse plane. Measure the area of
traverse plane, A3, which is located at the tip of the
Pitot-static tube.
3. Determine Ps2 for each fan by averaging the
pressure measurements at each of four static
pressure taps located in the short length of duct
between the outlet damper and the plenum. See
Annex E for details of static pressure taps. Measure
td2 in Plane 2 for each fan.
4. For each fan, measure td1 and tw1 in the path of the
air flowing into the fan inlet. Determine pb for the
general vicinity of the fans. Measure td3 in Plane 3. All
of these measurements are used in the determination
of densities at the various planes of interest.
5. Measure the fan speed and the motor amps, volts,
and if possible, watts for each fan. Record all
pertinent motor nameplate data, including volts
(NPV) and full load amps (FLA). If the motor power
outputs are to be estimated by using the phase
current method described in Annex K, it is not
necessary to measure motor watts; however, it may
be necessary to disconnect the drives and measure
the no load amps (NLA) if the motors are not
operating at or near their full load points. Refer to
Annex K.
6. SEF 1 is due to the effect of insufficient length of
duct between the outlet of each fan and the plenum.
In this case, the duct length is so short as to be
judged equivalent to there being no duct at all. In
order to calculate the value of SEF 1, it is necessary
to measure the outlet areas of the fans, A2, and their
blast areas.
Field Performance Measurement | 175
7. The sum of the static pressure, Ps1, and the
velocity pressure, Pv1, at the inlet of a fan with an
unrestricted inlet is considered to be equal to the sum
of the static pressure, Psx, and the velocity pressure,
Pvx, at a point sufficiently distant from the inlet as to
be in still air. At this point, the static pressure is zero,
and the velocity pressure in still air is zero.
Ps1 + Pv1 = Psx + Pvx
=0
This consideration, which is the same as that used in
the methods for testing fans for performance rating
purposes, charges to the fan losses incurred in
accelerating the air into the fan inlet and eliminates
the inaccuracies which arise in any attempt to
measure the velocity pressure and static pressure at
the fan inlet. To calculate the Fan Static Pressure:
Ps = Ps2 - Ps1 - Pv1 + SEF 1
= Ps2 - (Ps1 + Pv1) + SEF 1
Since Ps1 + Pv1 = 0:
Ps = Ps2 + SEF 1
8. In order to compare the test results to the quoted
fan curve drawn for operation at 1900 rpm and 0.075
lbm/ft3 density, it is necessary to convert the results
to the specified conditions. The basis for the
calculations is described in Section 14.
OBSERVATIONS
SITE MEASUREMENTS
pb = 29.05 in. Hg
td3 = 78°F
Ps3 = 5.6 in. wg
Pv3 = 0.47 in. wg
A3 = 7.4 ft2
LH Fan
td1 = 75°F
tw1 = 57°F
td2 = 79°F
Ps2 = 6.4 in. wg
N = 1910 rpm, LH fan speed
A2 = 3.2 ft2
Blast Area = 2.25 ft2
RH Fan
td1 = 75°F
tw1 = 57°F
td2 = 79°F
176 | Field Performance Measurement
Ps2 = 6.4 in. wg
N = 1890 rpm, RH fan speed
A2 = 3.2 ft2
Blast Area = 2.25 ft2
MEASURED MOTOR DATA
LH Fan
Volts = 575, 572, 578
= 575 av
Amps = 16, 17, 17
= 16.7 av
NLA = 7.0
RH Fan
Volts = 575, 574, 573
= 574 av
Amps = 15, 16, 16
= 15.7 av
NLA = 7.0
MOTOR NAMEPLATE DATA
LH Fan
25 hp, 3 phase, 60 hertz
575 volts, 1780 rpm, 23 FLA
RH Fan
25 hp, 3 phase, 60 hertz
575 volts, 1780 rpm, 23 FLA
GENERAL
Outlet dampers in full open positions. Fans
connected to motors through belt drives.
CALCULATIONS
DENSITIES
For inlet conditions for both fans of:
td1 = 75°F
tw1 = 57°F
p1 = pb
= 29.05 in. Hg
Use Figure N.1 in Annex N to obtain ρ1 = 0.0718
lbm/ft3
The density at Plane 2:
⎛ P + 13.6 pb ⎞ ⎛ t d1 + 460 ⎞
ρ2 = ρ1 ⎜ s2
⎟
⎟⎜
⎝ 13.6 p1 ⎠ ⎝ t d2 + 460 ⎠
⎛ 6.4 + 13.6 × 29.05 ⎞ ⎛ 535 ⎞
= 0.0718 ⎜
⎟⎜
⎟
⎝ 13.6 × 29.05 ⎠ ⎝ 539 ⎠
= 0.0724 lbm/ft 3
Annex K indicates that the average of the results of
Equation A and Equation B will provide a reasonably
accurate estimate of motor power output for a 25 hp
motor operating at approximately 70% FLA.
LH Fan
Eqn A = 25 (16.7/23) (575/575)
= 18.15 hp
The density at Plane 3:
⎛ P + 13.6 pb ⎞ ⎛ t d1 + 460 ⎞
ρ3 = ρ1 ⎜ s3
⎟
⎟⎜
⎝ 13.6 p1 ⎠ ⎝ t d3 + 460 ⎠
⎛ 5.6 + 13.6 × 29.05 ⎞ ⎛ 535 ⎞
= 0.0718 ⎜
⎟⎜
⎟
⎝ 13.6 × 29.05 ⎠ ⎝ 538 ⎠
= 0.0724 lbm/ft 3
Eqn B = 25 [(16.7 - 7)/(23 - 7)] (575/575)
= 15.16 hp
Hmo
= (18.15 + 15.16)/2
= 16.66 hp
RH Fan
Eqn A = 25 (15.7/23) (574/575)
= 17.04 hp
FLOW RATES
V3 = 1096 (Pv3/ρ3)0.5
= 1096 (0.47/0.0724)0.5
= 2792 fpm
Eqn B = 25 [(15.7 - 7)/(23 - 7)] (574/575)
= 13.57 hp
Hmo
= (17.04 + 13.57)/2
= 15.31 hp
Q3 = V3A3
= 2792 × 7.4
= 20661 cfm
Figure L.1 in Annex L indicates estimated belt drive
loss of 5% for each fan.
Q = Q1
= Q3 (ρ3/ρ1)
= 20661 (0.0724/0.0718)
= 20834 cfm
Assume that the air flow rate for each fan is
proportional to its speed.
LH Motor
HL = 0.05 Hmo
= 0.05 × 16.66
= 0.83 hp
H = Hmo - HL
= 16.66 - 0.83
= 15.83 hp
LH Fan
Q = Q1
= 20834 [1910/(1910 + 1890)]
= 10472 cfm
RH Fan
Q = Q1
= 20834 [1890/(1910 + 1890)]
= 10362 cfm
FAN POWER INPUT
LH Fan
Measured amps/FLA = (16.7/23)
= 0.73
= 73%
RH Fan
Measured amps/FLA = (15.7/23)
= 0.68
= 68%
RH Motor
HL = 0.05 Hmo
= 0.05 × 15.31
= 0.77 hp
H = Hmo - HL
= 15.31 - 0.77
= 14.54 hp
SYSTEM EFFECT FACTOR
AMCA Publication 201-90, Figures 7.1 and 8.3
indicate the following calculations:
LH Fan
Q2 = Q1 (ρ1/ρ2)
= 10472 (0.0718/0.0724)
= 10385 cfm
V2 = (Q2/A2)
= (10385/3.2)
= 3245 fpm
Field Performance Measurement | 177
Blast area ratio = Blast Area/A2
= 2.25/3.2
= 0.70
RH Fan
Q2 = Q1 (ρ1/ρ2)
= 10362 (0.0718/0.0724)
= 10276 cfm
V2 = (Q2/A2)
= (10276/3.2)
= 3211 fpm
Blast area ratio = Blast Area/A2
= 2.25/3.2
= 0.70
For a blast area ratio of 0.7 and no duct, Figure 8.3
shows System Effect Curve S applies. For each fan
with velocities of 3245 fpm and 3211 fpm and curve
S, Figure 7.1 shows SEF 1 = 0.5 in. wg at 0.075
lbm/ft3. At 0.0724 lbm/ft3:
SEF 1 = 0.5 (0.0724/0.075)
= 0.48 in. wg
FAN STATIC PRESSURE
Ps = Ps2 + SEF 1
LH Fan
Ps = 6.4 + 0.48
= 6.88 in. wg
RH Fan
Ps = 6.4 + 0.48
= 6.88 in. wg
178 | Field Performance Measurement
CONVERSION TO SPECIFIED CONDITIONS
LH Fan
Qc = 10472 (1900/1910)
= 10417 cfm
Psc = 6.88 (1900/1910)2 (0.075/0.0718)
= 7.11 in. wg
Hc = 15.83 (1900/1910)3 (0.075/0.0718)
= 16.28 hp
RH Fan
Qc = 10362 (1900/1890)
= 10417 cfm
Psc = 6.88 (1900/1890)2 (0.075/0.0718)
= 7.26 in. wg
Hc = 14.54 (1900/1890)3 (0.075/0.0718)
= 15.43 hp
EXAMPLE 2A: UTILITY FAN IN A VENTILATION SYSTEM
3
STATIC PRESSURE TAPS
1
2
PLAN VIEW
L
3-PIECE
ELBOW
R/D = 1
SEF 1
SEF 2
SIDE VIEW
OUTLET SIDE VIEW
COMMENTS
1. Determine Pv3 by using the root mean square of
the velocity pressure measurements made in a
traverse of Plane 3, located near the end of a straight
run of duct, such as shown in the diagram. Determine
Ps3 by averaging the static pressure measurements
made in the same traverse. Procedures for traverses
are described in Section 9.4. Ps3 is used in
determining the density at the traverse plane.
Measure the area of the traverse plane, A3, which is
located at the tip of the Pitot-static tube.
2. Determine Ps1 by averaging the pressure
measurements at each of four static pressure taps in
the collar connection at the fan inlet. Determine Ps2
by averaging the pressure measurements at each of
four static pressure taps located near the fan outlet.
3. Measure td3 and tw3 in the traverse plane. Assume
td1 is equal to td3. Determine pb for the general vicinity
of the fan. Measure td2 in Plane 2. All of these
measurements are used in determining densities at
the various planes of interest.
4. Measure the fan speed and the motor amps, volts,
and if possible, watts. Record all pertinent motor
nameplate data, including volts (NPV) and full load
amps (FLA). If the motor power output is to be
estimated by using the phase current method
described in Annex K, it is not necessary to measure
motor watts; however, it may be necessary to
disconnect the drive and measure the no load amps
(NLA) if the motor is not operating at or near its full
load point. Refer to Annex K.
5. SEF 1 is due to the effect of the elbow located at
the fan inlet. SEF 2 is due to the effect of insufficient
length of duct between the fan outlet and the elbow
downstream of the fan. In order to calculate the
values of the SEFs, it is necessary to measure the
inlet area and the outlet area of the fan, A1 and A2;
the length of the outlet duct, L; and the blast area of
the fan.
6. To calculate the Fan Static Pressure:
Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2
Where:
Pv1 = Pv3 (A3/A1)2 (ρ3/ρ1)
7. In order to compare the test results to the quoted
fan curve drawn for operation at 1880 rpm and 0.075
lbm/ft3 density, it is necessary to convert the results
to the specified conditions. The basis for the
calculations is described in Section 14.
Field Performance Measurement | 179
OBSERVATIONS
SITE MEASUREMENTS
pb = 29.20 in. Hg
td2 = 72°F
td3 = 72°F
tw3 = 66°F
Ps1 = -2.18 in. wg
Ps2 = 0.35 in. wg
Ps3 = -1.95 in. wg
Pv3 = 0.45 in. wg
N = 1730 rpm
A1 = 1.07 ft2
A2 = 1.17 ft2
A3 = 1.07 ft2
Blast Area = 0.7 ft2
L = 0.83 ft
MEASURED MOTOR DATA
Volts = 227, 229, 228
= 228 av
Amps = 10.2, 10.3, 10.4
= 10.3 av
NLA = 7.1
MOTOR NAMEPLATE DATA
5 hp, 3 phase, 60 hertz
230 volts, 1750 rpm, 14 FLA
The density at Plane 1:
⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞
ρ1 = ρ3 ⎜ s1
⎟⎜
⎟
⎝ 13.6 p3 ⎠ ⎝ t d1 + 460 ⎠
⎛ −2.18 + 13.6 × 29.20 ⎞ ⎛ 532 ⎞
= 0.0719 ⎜
⎟ ⎜ 532 ⎟
13.6 × 29.06
⎝
⎠⎝
⎠
= 0.0718 lbm/ft 3
The density at Plane 2:
⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞
ρ2 = ρ3 ⎜ s2
⎟⎜
⎟
⎝ 13.6 p3 ⎠ ⎝ t d2 + 460 ⎠
⎛ 0.35 + 13.6 × 29.20 ⎞ ⎛ 532 ⎞
= 0.0719 ⎜
⎟ ⎜ 532 ⎟
13.6 × 29.06
⎝
⎠⎝
⎠
= 0.0723 lbm/ft 3
FLOW RATES
V3 = 1096 (Pv3/ρ3)0.5
= 1096 (0.45/0.0719)0.5
= 2742 fpm
Q3 = V3A3
= 2742 × 1.07
= 2934 cfm
GENERAL
Q = Q1
= Q3 (ρ3/ρ1)
= 2934 (0.0719/0.0718)
= 2938 cfm
Fan connected to motor through belt drive.
FAN POWER INPUT
CALCULATIONS
DENSITIES
For Plane 3 conditions of:
td3 = 72°F
tw3 = 66°F
p3 = pb + (Ps3/13.6)
= 29.20 + (-1.95/13.6)
= 29.06 in. Hg
Use Figure N.1 in Annex N to obtain ρ3 = 0.0719
lbm/ft3
Measured amps/FLA = 10.3/14
= 0.74
= 74%
Annex K indicates that the average of the results of
Equation A and Equation B will provide a reasonably
accurate estimate of motor power output for a 5 hp
motor operating at 74% FLA.
Eqn A = 5 (10.3/14) (228/230)
= 3.65 hp
Eqn B = 5 [(10.3 - 7.1)/(14 - 7.1)] (228/230)
= 2.30 hp
Hmo
It is assumed that td1 = td3
180 | Field Performance Measurement
= (3.65 + 2.30)/2
= 2.98 hp
Figure L.1 in Annex L indicates estimated belt drive
loss of 6.5%.
HL = 0.065 Hmo
= 0.065 × 2.98
= 0.19 hp
H = Hmo - HL
= 2.98 - 0.19
= 2.79 hp
SYSTEM EFFECT FACTORS
To determine the value of SEF 1, calculate the
velocity at the fan inlet:
V1 = Q1/A1
= 2938/1.07
= 2746 fpm
L in % effective duct length
= (L/3.05) 100
= (0.83/3.05) 100
= 27%
Blast area ratio = Blast Area/A2
= 0.7/1.17
= 0.6
For blast area ratio of 0.6, 27% effective duct length
and elbow position C, Figure 8.5 shows System
Effect Curve P - Q applies. For 2494 fpm velocity and
curve P - Q, Figure 7.1 shows SEF 2 = 0.7 in. wg at
0.075 lbm/ft3. At 0.0723 lbm/ft3:
SEF 2 = 0.7 (0.0723/0.075)
= 0.67 in. wg
FAN STATIC PRESSURE
AMCA Publication 201-90, Figure 9.5 indicates that
for a three piece elbow with radius to diameter ratio
of 1, and with no duct between the elbow and the fan
inlet, System Effect Curve R applies. For 2746 fpm
velocity and curve R, Figure 7.1 shows SEF 1 = 0.55
in. wg at 0.075 lbm/ft3. At 0.0718 lbm/ft3:
SEF 1 = 0.55 (0.0718/0.075)
= 0.53 in. wg
Pv1 = Pv3 (A3/A1)2 (ρ3/ρ1)
= 0.45 (1.07/1.07)2 (0.0719/0.0718)
= 0.45 in. wg
Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2
= 0.35 - (-2.18) - 0.45 + 0.53 + 0.67
= 3.28 in. wg
CONVERSION TO SPECIFIED CONDITIONS
For SEF 2, AMCA Publication 201-90, Figures 7.1,
8.1, and 8.5 indicate the following calculations:
Q2 = Q3 (ρ3/ρ2)
= 2934 (0.0719/0.0723)
= 2918 cfm
V2 = (Q2/A2)
= 2918/1.17
= 2494 fpm
Qc = 2938 (1880/1730)
= 3193 cfm
Psc = 3.28 (1880/1730)2 (0.075/0.0718)
= 4.05 in. wg
Hc = 2.79 (1880/1730)3 (0.075/0.0718)
= 3.74 hp
Duct diameter equivalent to the fan outlet area:
De2 = (4A2/π)0.5
= (4 × 1.17/π)0.5
= 1.22 ft
Figure 8.1 shows that for velocities of 2500 fpm or
less, the 100% effective outlet duct length is 2.5 duct
diameters,
= 2.5 × 1.22
= 3.05 ft
Field Performance Measurement | 181
EXAMPLE 2B: CENTRIFUGAL FAN IN A SAWDUST CONVEYING SYSTEM
SEF 2 2
1
SEF 1
4-PIECE ELBOW
R/D = 1
L2
L1
3
SIDE VIEW
OUTLET SIDE VIEW
COMMENTS
1. Determine Pv3 by using the root mean square of
the velocity pressure measurements made in a
traverse of Plane 3, located near the end of a straight
run of duct, such as shown in the diagram. Determine
Ps3 by averaging the static pressure measurements
made in the same traverse. Procedures for traverses
are described in Section 9.4. Ps3 is used in
determining the density at the traverse plane.
Measure the area of the traverse plane, A3, which is
located at the tip of the Pitot-static tube.
2. Determine Ps1 by using a Pitot-static tube or static
pressure taps in the duct connection at the fan inlet.
If a Pitot-static tube is used, it should not project into
the upstream elbow but be located well within the
length of the duct connection as shown in the
diagram. The friction loss in the short length of outlet
duct is assumed to be negligible, and Ps2 is
considered to be equal to the static pressure at the
duct outlet. The static pressure at the outlet of the
duct is zero gauge pressure, referred to the
atmospheric pressure in the region of the duct outlet.
In situations such as this example, the air may be
discharging from the duct into a region in which the
atmospheric pressure is somewhat different from that
to which all other pressure measurements are
referred. When this possibility exists, it is essential
that the static pressure in the region of the
discharging air be measured, referred to the same
182 | Field Performance Measurement
atmospheric pressure as used in all other pressure
measurements. In this case, the pressure was
measured as 0.1 in. wg.
3. Measure td3 and tw3 in the traverse plane.
Determine pb for the general vicinity of the fan.
Measure td1 and td2. All of these measurements are
used in determining densities at the various planes of
interest.
4. Measure the fan speed and the motor amps, volts,
and if possible, watts. Record all pertinent motor
nameplate data, including volts (NPV) and full load
amps (FLA). If the motor power output is to be
estimated by using the phase current method
described in Annex K, it is not necessary to measure
motor watts; however, it may be necessary to
disconnect the drive and measure the no load amps
(NLA) if the motor is not operating at or near its full
load point. Refer to Annex K.
5. SEF 1 is due to the effect of insufficient length of
duct between the fan inlet and the elbow upstream of
the fan. SEF 2 is due to the effect of insufficient
length of duct at the fan outlet. In order to calculate
the values of the SEFs, it is necessary to measure
the inlet area and the outlet area of the fan, A1 and
A2; the lengths of the inlet connection and the outlet
duct, L1 and L2; and the blast area of the fan.
6. To calculate the Fan Static Pressure:
CALCULATION
Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2
DENSITIES
Where:
For Plane 2 conditions of:
Pv1 = Pv3 (A3/A1)2 (ρ3/ρ1)
td2 = 91.3°F
tw2 = 70.4°F
7. In order to compare the test results to the quoted
fan curve drawn for operation at 2075 rpm and 0.075
lbm/ft3, it is necessary to convert the results to the
specified conditions. The basis for the calculations is
described in Section 14.
OBSERVATIONS
p2 = pb + (Ps2/13.6)
= 29.82 + (0.1/13.6)
= 29.83 in. Hg
Use Figure N.1 in Annex N to obtain ρ2 = 0.0714
lbm/ft3
SITE MEASUREMENTS
The density at Plane 1:
pb = 29.82 in. Hg
td1 = 86.6°F
td2 = 91.3°F
tw2 = 70.4°F
td3 = 86°F
Ps1 = -11.4 in. wg
Ps2 = 0.1 in. wg
Ps3 = -8.9 in. wg
Pv3 = 1.24 in. wg
N = 2120 rpm, fan speed
A1 = 1.40 ft2
A2 = 1.40 ft2
A3 = 1.57 ft2
⎛ P + 13.6 pb ⎞ ⎛ t d2 + 460 ⎞
ρ1 = ρ2 ⎜ s1
⎟
⎟⎜
⎝ 13.6 p2 ⎠ ⎝ t d1 + 460 ⎠
4 + 13.6 × 29.82 ⎞ ⎛ 551.3 ⎞
⎛ −11.4
= 0.0714 ⎜
⎟ ⎜ 546.6 ⎟
13.6 × 29.83
⎝
⎠⎝
⎠
Blast Area = 1.26 ft2
L1 = 1.33 ft
L2 = 3.0 ft
MEASURED MOTOR DATA
Volts = 460, 460, 459
= 460 av
Amps = 26.5, 25.5, 26
= 26 av
NLA = 11.3
= 0.0700 lbm/ft 3
The density at Plane 3:
⎛ P + 13.6 pb ⎞ ⎛ t d2 + 460 ⎞
ρ3 = ρ2 ⎜ s3
⎟⎜
⎟
⎝ 13.6 p3 ⎠ ⎝ t d3 + 460 ⎠
⎛ −8.9 + 13.6 × 29.82 ⎞ ⎛ 551.3 ⎞
= 0.0714 ⎜
⎟ ⎜ 546 ⎟
13.6 × 29.83
⎝
⎠⎝
⎠
= 0.0705 lbm/ft 3
FLOW RATES
V3 = 1096 (Pv3/ρ3)0.5
= 1096 (1.24/0.0705)0.5
= 4596 fpm
Q3 = V3A3
= 4596 × 1.57
= 7216 cfm
MOTOR NAMEPLATE DATA
30 hp, 3 phase, 60 hertz
460 volts, 1750 rpm, 36 FLA
GENERAL
Fan connected to motor through belt drive.
Q = Q1
= Q3 (ρ3/ρ1)
= 7216 (0.0705/0.0700)
= 7268 cfm
FAN POWER INPUT
Measured amps/FLA = (26/36)
= 0.72
= 72%
Field Performance Measurement | 183
Annex K indicates that the average of the results of
Equation A and Equation B will provide a reasonably
accurate estimate of motor power output for a 30 hp
motor operating at 72% FLA.
Eqn A = 30 (26/36) (460/460)
= 21.67 hp
Eqn B = 30 [(26 - 11.3)/(36 - 11.3)] (460/460)
= 17.85 hp
Hmo
= (21.67 + 17.85)/2
= 19.76 hp
For SEF 2, AMCA Publication 201-90, Figure 8.3
indicates the following calculations:
Q2 = Q3 (ρ3/ρ2)
= 7216 (0.0705/0.0714)
= 7125 cfm
V2 = (Q2/A2)
= (7125/1.40)
= 5089 fpm
Duct diameter equivalent to the fan outlet area:
Figure L.1 in Annex L indicates estimated belt drive
loss of 4.8%.
De2 = (4A2/π)0.5
= (4 × 1.40/π)0.5
= 1.34 ft
HL = 0.048 Hmo
= 0.048 × 19.76
= 0.95 hp
Figure 8.3 shows that for velocities over 2500 fpm,
100% effective duct length is one duct diameter per
1000 fpm:
H = Hmo - HL
= 19.76 - 0.95
= 18.81 hp
SYSTEM EFFECT FACTORS
To determine the value of SEF 1, calculate the
velocity at the fan inlet:
V1 = (Q1/A1)
= (7268/1.40)
= 5191 fpm
= D2 (V2/1000)
= 1.34 (5089/1000)
= 6.82 ft
The length of the outlet duct in % effective duct
length:
= (L2/6.82) 100
= (3.0/6.82) 100
= 44%
The diameter of the fan inlet:
Blast ratio area = Blast Area/A2
= 1.26/1.40
= 0.9
D1 = (4A1/π)0.5
= (4 × 1.40/π)0.5
= 1.34 ft.
For blast area ratio of 0.9 and 44% effective duct
length, Figure 8.3 shows no System Effect Curve
applies and SEF 2 = 0.
The length of the duct between the elbow and the fan
inlet in terms of D1:
FAN STATIC PRESSURE
= (L1/D1)
= (1.33/1.34)
= 1.0
AMCA Publication 201-90, Figure 9.5 indicates that
for a four piece elbow with a radius to diameter ratio
of 1, and with a length of duct between the elbow and
the fan inlet equal to 1 equivalent diameter, System
Effect Curve S applies. For 5191 fpm velocity and
curve S, Figure 7.1 shows SEF 1 = 1.3 in. wg at
0.075 lbm/ft3. At 0.0700 lbm/ft3:
SEF 1 = 1.3 (0.0700/0.075)
= 1.2 in. wg
184 | Field Performance Measurement
Pv1 = Pv3 (A3/A1)2 (ρ3/ρ1)
= 1.24 (1.57/1.40)2 (0.0705/0.0700)
= 1.57 in. wg
Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2
= 0.1 - (-11.4) - 1.57 + 1.2 + 0
= 11.13 in. wg
CONVERSIONS TO SPECIFIED CONDITIONS
Qc = 7268 (2075/2120)
= 7114 cfm
Psc = 11.13 (2075/2120)2 (0.075/0.0700)
= 11.42 in. wg
Hc = 18.81 (2075/2120)2 (0.075/0.0700)
= 18.90 hp
Field Performance Measurement | 185
EXAMPLE 2C: AXIAL FAN IN A DRYER SYSTEM
5
4
1
STRAIGHTENING VANES
3
2
SEF 2
STATIC PRESSURE TAPS
A3
SEF 1
PLAN VIEW
INNER CYLINDER
LOCATION OF PLANE 3
SIDE VIEW
COMMENTS
1. This type of installation is normally classified as
one in which a satisfactory test cannot be conducted.
Due to the configurations of the airways, there are no
locations at which reasonably accurate pressure
measurements can be made. In addition, the
judgments required in determining the values of the
SEFs are susceptible to error. The purpose of
presenting this example is to illustrate the not
uncommon instance in which a test must be conducted
in order to provide performance information, even
though the results will be innaccurate to a degree
which is not normally acceptable.
2. Determine Pv3 by using the root mean square of
the velocity pressure measurements made in a
traverse of Plane 3, located as shown in the diagram.
Determine Ps3 by averaging the static pressure
measurements made in the same traverse.
Procedures for traverses are described in Section
9.4. These velocity pressure and static pressure
measurements are susceptible to error due to the
turbulence existing in the region of the fan outlet. In
addition, it is undesirable to have Plane 3 located in
a diverging airway. However, no other more suitable
location for Plane 3 exists in this example. It is
recommended that the Pitot-static tube be oriented
so that its nose is aligned with the anticipated flow
streams, particularly near the walls of the diffuser.
Determine the area of the traverse plane, A3, which is
186 | Field Performance Measurement
located at the tip of the Pitot-static tube, as shown in
the diagram, not at the location of the Pitot-static tube
access holes.
3. Determine Ps4 by averaging the pressure
measurements at each of four static pressure taps
located near the fan inlet. In the same manner,
determine Ps5 at a location near the fan outlet. It is
undesirable to have pressure measurement planes
located in converging and diverging airways, but
there are no other more suitable locations for these
planes in this installation. Measure A4 and A5, the
cross-sectional areas of the airways at Planes 4 and 5.
4. Measure td3, tw3, and td4. Determine pb for the
general vicinity of the fan. These measurements are
used in the determination of densities at the various
planes of interest.
5. Measure the fan speed and the motor amps, volts,
and if possible, watts. Record all pertinent motor
nameplate data, including volts (NPV) and full load
amps (FLA). If the motor power output is to be
estimated by using the phase current method
described in Annex K, it is not necessary to measure
motor watts; however, it may be necessary to
disconnect the drive and measure the no load amps
(NLA) if the motor is not operating at or near its full
load point. Refer to Annex K.
6. Although an elbow is located a short distance
upstream of the fan, it is considered to produce no
system effect since it is equipped with turning vanes
and the average velocity through the elbow will be
relatively low due to its large cross-sectional area.
Therefore, SEF 1 = 0. In judging SEF 2, the rapidly
diverging transition fitting downstream of the fan is
considered equivalent to no duct at the fan outlet. In
order to calculate the value of SEF2, it is necessary
to measure the outlet area of the fan, A2.
MOTOR NAMEPLATE DATA
7. To calculate the Fan Static Pressure,
DENSITIES
Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2
For Plane 3 conditions of:
Where:
td3 = 86.5°F
tw3 = 75.5°F
25 hp, 3 phase, 60 hertz
460 volts, 1750 rpm, 31 FLA
GENERAL
Fan connected to motor through belt drive
CALCULATIONS
Pv1 = Pv3 (A3/A1)2 (ρ3/ρ1)
Ps1 and Ps2 are calculated on the basis of total
pressure considerations, using Ps4, Ps5, and the
calculated velocity pressures at Planes 1, 2, 4, and 5.
8. In order to compare the test results to the quoted
fan curve drawn for operation at 1580 rpm and
0.0690 lbm/ft3 density, it is necessary to convert the
results to the specified conditions. The basis for the
calculations is described in Section 14.
OBSERVATIONS
SITE MEASUREMENTS
pb = 28.90 in. Hg
td3 = 86.5°F
tw3 = 75.5°F
td4 = 85°F
Ps3 = 1.5 in. wg
Pv3 = 0.044 in. wg
Ps4 = -1.57 in. wg
Ps5 = 1.22 in. wg
N = 1590 rpm
A1 = A2 = 8.0 ft2
A3 = 29.8 ft2
A4 = 12.4 ft2
A5 = 9.6 ft2
MEASURED MOTOR DATA
Volts = 450, 449, 448
= 449 av
Amps = 25.0, 24.5, 25.0
= 24.8 av
NLA = 9.4
p3 = pb + (Ps3/13.6)
= 28.90 + (1.5/13.6)
= 29.01 in. Hg
Use Figure N.1 from Annex N to obtain ρ3 = 0.0694
lbm/ft3
The density at Plane 4:
⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞
ρ 4 = ρ3 ⎜ s4
⎟⎜
⎟
⎝ 13.6 p3 ⎠ ⎝ t d4 + 460 ⎠
⎛ −1.57 + 13.6 × 28.90 ⎞ ⎛ 546.5 ⎞
= 0.0694 ⎜
⎟ ⎜ 545 ⎟
13.6 × 29.01
⎝
⎠⎝
⎠
= 0.0691 lbm/ft 3
It is assumed that td1 = td4 and at the low pressure
levels which exist at Planes 1 and 4, the difference
between these pressures will be small, and assuming
ρ1 = ρ4, will result in an error which is considered
negligible. By similar reasoning, it is assumed that
ρ5 = ρ2 = ρ3.
FLOW RATES
V3 = 1096 (Pv3/ρ3)0.5
= 1096 (0.044/0.0694)0.5
= 873 fpm
Q3 = V3A3
= 873 × 29.8
= 26015 cfm
Q = Q1
= Q3 (ρ3/ρ1)
= 26015 (0.0694/0.0691)
= 26128 cfm
Field Performance Measurement | 187
FAN POWER INPUT
FAN STATIC PRESSURE
Measured amps/FLA = (24.8/31)
= 0.80
= 80%
Pv4 = Pv3 (A3/A4)2 (ρ3/ρ4)
= 0.044 (29.8/12.4)2 (0.0694/0.0691)
= 0.26 in. wg
Annex K indicates that the average of the results of
Equation A and Equation B will provide a reasonably
accurate estimate of motor power output for a 25 hp
motor operating at 80% FLA.
Pv1 = Pv3 (A3/A1)2 (ρ3/ρ1)
= 0.044 (29.8/8.0)2 (0.0694/0.0691)
= 0.61 in. wg
Eqn A = 25 (24.8/31) (449/460)
= 19.52 hp
Ps1 + Pv1 = Ps4 + Pv4
Ps1 = Ps4 + Pv4 - Pv1
= -1.57 + 0.26 - 0.61
= -1.92 in. wg
Eqn B = 25 [(24.8 - 9.4)/(31 - 9.4)] (449/460)
= 17.40 hp
Hmo
= (19.52 + 17.40)/2
= 18.46 hp
Figure L.1 in Annex L indicates estimated belt drive
loss of 4.9%.
HL = 0.049 Hmo
= 0.049 × 18.46
= 0.90 hp
H = Hmo - HL
= 18.46 - 0.90
= 17.56 hp
SYSTEM EFFECT FACTORS
SEF 1 = 0 See item 6 under COMMENTS.
To determine the value of SEF 2, AMCA Publication
201-90, Figure 8.2 indicates that a vaneaxial fan with
no outlet duct will use System Effect Curve U.
Q2 = Q3 (ρ3/ρ2)
= 26015 (0.0694/0.0694)
= 26015 cfm
V2 = (Q2/A2)
= (26015/8.0)
= 3252 fpm
From Figure 7.1, using 3252 fpm and curve U, SEF 2
= 0.26 in. wg at 0.075 lbm/ft3. At 0.0694 lbm/ft3:
SEF 2 = 0.26 (0.0694/0.075)
= 0.24 in. wg
188 | Field Performance Measurement
Pv5 = Pv3 (A3/A5)2 (ρ3/ρ5)
= 0.044 (29.8/9.6)2 (0.0694/0.0694)
= 0.42 in. wg
Pv2 = Pv3 (A3/A2)2 (ρ3/ρ2)
= 0.044 (29.8/8.0)2 (0.0694/0.0694)
= 0.61 in. wg
Ps2 + Pv2 = Ps5 + Pv5
Ps2 = Ps5 + Pv5 - Pv2
= 1.22 + 0.42 - 0.61
= 1.03 in. wg
Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2
= 1.03 - (-1.92) - 0.61 + 0 + 0.24
= 2.58 in. wg
Losses between Planes 1 and 4 and between Planes
2 and 5 have been ignored.
CONVERSION TO SPECIFIED CONDITIONS
Qc = 26128 (1580/1590)
= 25964 cfm
Psc = 2.58 (1580/1590)2 (0.0690/0.0691)
= 2.54 in. wg
Hc = 17.56 (1580/1590)3 (0.0690/0.0691)
= 17.21 hp
EXAMPLE 2D: CENTRIFUGAL FAN IN A SCRUBBER SYSTEM
INLET BOX DAMPER
STATIC PRESSURE TAPS
SEF 1
3
1
L
INLET BOX
2
DIFFUSER
SIDE VIEW
OUTLET SIDE VIEW
COMMENTS
1. This fan, as supplied and rated by the
manufacturer, includes the inlet box damper and the
inlet box. Performance ratings for fans with inlet box
dampers cover operation with the dampers in the full
open positions. In order to be able to compare the
test results to the fan performance ratings, it is
essential that the damper be fixed in the full open
position for the duration of the test.
2. Determine Pv3 by using the root mean square of
the velocity pressure measurements made in a
traverse of Plane 3, located shortly upstream of the
inlet damper. Determine Ps3 by averaging the static
pressure measurements made in the same traverse.
Procedures for traverses are described in Section
9.4. Measure A3, the area of the traverse plane,
located at the tip of the Pitot-static tube and A1, the
area of the inlet to the damper.
3. Determine Ps2 by averaging the pressure
measurements at each of four static pressure taps
located near the end of the fan outlet. See Annex E
for details of static pressure taps.
4. Measure td3 and tw3 in the traverse plane.
Determine pb for the general vicinity of the fan.
Measure td2 in Plane 2. These measurements are
used in the determination of densities at the various
planes of interest.
5. Measure the fan speed and the motor amps, volts,
and if possible, watts. Record all pertinent motor
nameplate data, including volts (NPV), and full load
amps (FLA). If the motor power output is to be
estimated by using the phase current method
described in Annex K, it is not necessary to measure
motor watts; however, it may be necessary to
disconnect the drive and measure the no load amps
(NLA) if the motor is not operating at or near its full
load point. Refer to Annex K.
6. SEF 1 is due to the effect of insufficient length of
duct at the fan outlet. In order to calculate the value
of SEF 1, it is necessary to measure the length of the
outlet duct, L; the fan outlet area, A2; and the blast
area of the fan.
7. To calculate the Fan Static Pressure:
Ps = Ps2 - Ps1 - Pv1 + SEF 1
Since Plane 1 is located shortly downstream of Plane
3 in an airway of uniform cross-section (A1 = A3), the
conditions which exist at Plane 3 are assumed to
exist at Plane 1. Therefore, Ps1 = Ps3 and Pv1 = Pv3.
8. In order to compare the test results to the quoted
fan curve drawn for operation at 1780 rpm and 0.059
lbm/ft3 density, it is necessary to convert the results
Field Performance Measurement | 189
to the specified conditions. The basis for the
calculations is described in Section 14.
OBSERVATIONS
pe = 0.5603 in. Hg
pp = pe - [p3 (td3 - tw3)/2700]
= 0.5603 - [24.28 (63 - 62)/2700]
= 0.5513 in. Hg
SITE MEASUREMENTS
pb = 29.44 in. Hg
td2 = 97°F
td3 = 63°F
tw3 = 62°F
Ps2 = 1.1 in. wg
Ps3 = -70.2 in. wg
Pv3 = 0.64 in. wg
N = 1790 rpm
A1 = 6.5 ft2
A2 = 5.32 ft2
A3 = 6.5 ft2
Blast Area = 1.89 ft2
L = 2.50 ft
MEASURED MOTOR DATA
Volts = 4160, 4150, 4150
= 4153 av
Amps = 50, 51, 52
= 51 av
NLA = 14
MOTOR NAMEPLATE DATA
500 hp, 3 phase, 60 hertz
4160 volts, 1785 rpm, 61 FLA
GENERAL
Inlet box damper in full open position. Fan direct
connected to motor.
CALCULATIONS
DENSITIES
For Plane 3 conditions of:
td3 = 63°F
tw3 = 62°F
p3 = pb + (Ps3/13.6)
= 29.44 + (-70.2/13.6)
= 24.28 in. Hg
Use the modified Apjohn equation, described in
Section M.2.3 in Annex M, and the table in Figure N.2
in Annex N to calculate the density at Plane 3.
190 | Field Performance Measurement
ρ3 =
=
1.3257( p3 − 0.378 pp )
t d3 + 460
1.3257 ( 24.28 − 0.378 × 0.5513 )
63 + 460
= 0.0610 lbm/ft 3
Consider ρ1 to be equal to ρ3.
The density at Plane 2:
⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞
ρ2 = ρ3 ⎜ s2
⎟⎜
⎟
⎝ 13.6 p3 ⎠ ⎝ t d2 + 460 ⎠
⎛ 1.1 + 13.6 × 29.44 ⎞ ⎛ 523 ⎞
= 0.0610 ⎜
⎟⎜
⎟
⎝ 13.6 × 24.28 ⎠ ⎝ 557 ⎠
= 0.0696 lbm/ft 3
FLOW RATES
V3 = 1096 (Pv3/ρ3)0.5
= 1096 (0.64/0.0610)0.5
= 3550 fpm
Q3 = V3A3
= 3550 × 6.5
= 23075 cfm
Q = Q1 = Q3
= 23075 cfm
FAN POWER INPUT
Measured amps/FLA = 51/61
= 0.84
= 84%
Annex K indicates that the average of the results of
Equation A and Equation B will provide a reasonably
accurate estimate of motor power output for a 500 hp
motor operating at 84% FLA.
Eqn A = 500 (51/61) (4153/4160)
= 417 hp
Eqn B = 500 [(51 - 14)/(61 - 14)] (4153/4160)
= 393 hp
Hmo
= (417 + 393)/2
= 405 hp
Since the fan is direct-connected to the motor, there
is no drive loss, and:
H = Hmo
= 405 hp
SYSTEM EFFECT FACTOR
AMCA Publication 201-90, Figures 7.1 and 8.3
indicate the following calculations.
Q2 = Q3 (ρ3/ρ2)
= 23075 (0.0610/0.0696)
= 20224 cfm
V2 = Q2/A2
= 20224/5.32
= 3802 fpm
Duct diameter equivalent to the diffuser outlet area:
De2 = (4A2/π)0.5
= (4 × 5.32/π)0.5
= 2.60 ft
Figure 8.3 shows that for velocities over 2500 fpm
100% effective duct length is one duct diameter for
every 1000 fpm:
= De2 (V2/1000)
= 2.60 (3802/1000)
= 9.89 ft.
L in % effective duct length:
Blast area ratio = Blast Area/A2
= 1.89/5.32
= 0.36
For a blast area ratio of 0.36, and 25% effective duct
length, Figure 8.3 shows System Effect Curve U
applies. For 3802 fpm velocity and curve U, Figure
7.1 shows SEF 1 = 0.36 in. wg at 0.075 lbm/ft3. At
0.0696 lbm/ft3:
SEF 1 = 0.36 (0.0696/0.075)
= 0.33 in. wg
FAN STATIC PRESSURE
Ps1 = Ps3
= - 70.2 in. wg
Pv1 = Pv3
= 0.64 in. wg
Ps = Ps2 - Ps1 - Pv1 + SEF 1
= 1.1 - (-70.2) - 0.64 + 0.33
= 71.0 in. wg
CONVERSION TO SPECIFIED CONDITIONS
Qc = 23075 (1780/1790)
= 22946 cfm
Psc = 71.0 (1780/1790)2 (0.059/0.0610)
= 67.9 in. wg
Hc = 405 (1780/1790)3 (0.059/0.0610)
= 385 hp
= (L/9.89) 100
= (2.50/9.89) 100
= 25%
Field Performance Measurement | 191
EXAMPLE 2E: CENTRIFUGAL FAN IN A PROCESS SYSTEM
STATIC
OUTLET DAMPER
PRESSURE TAPS 5
2
INLET BOXES
INLET BOX
DAMPERS
1a
1b
3a
3b
OPPOSITE OUTLET SIDE VIEW
SIDE VIEW
COMMENTS
1. This fan, as supplied and rated by the
manufacturer, includes the inlet box dampers and the
inlet boxes, but does not include the outlet damper.
Performance ratings for fans with inlet box dampers
cover operation with the dampers in the full open
positions. Also, performance ratings for items such as
the outlet damper are for operation in the full open
position. In order to be able to compare the test
results to the fan performance ratings, it is essential
that the outlet damper and the inlet dampers be fixed
in their full open positions.
2. Determine Pv3a and Pv3b by using the root mean
square of the velocity pressure measurements made
in Planes 3a and 3b. Determine Ps3a and Ps3b by
averaging each of the two sets of static pressure
measurements made in the same traverses.
Procedures for traverses are described in Section
9.4. Measure A3a and A3b, the areas of the traverse
planes and A1a and A1b, the areas of the inlets to the
inlet dampers.
3. Determine Ps5 by averaging the pressure
measurements of each of four static pressure taps
located downstream of the outlet damper.
192 | Field Performance Measurement
4. Measure td3a, td3b, and td5. Since flue gas is being
handled by the fan, the Orsat apparatus is used by
process personnel to determine the density of the
gas. Determine pb for the general vicinity of the fan.
These data are used in the determination of densities
at the various planes of interest.
5. Measure the fan speed and the motor amps, volts,
and if possible, watts. Record all pertinent motor
nameplate data, including volts (NPV) and full load
amps (FLA). If the motor power output is to be
estimated by using the phase current method
described in Annex K, it is not necessary to measure
motor watts; however, it may be necessary to
disconnect the drive and measure the no load amps
(NLA) if the motor is not operating at or near its full
load point. Motor performance data, supplied by the
motor manufacturer, are used in the determination of
motor power output for this example.
6. In this example, the duct downstream of the outlet
damper is of sufficient length, and no SEF applies.
7. To calculate the Fan Static Pressure:
Ps = Ps2 - Ps1 - Pv1
Ps2 is calculated on the basis of total pressure
considerations using Ps5, the outlet damper pressure
loss, and the calculated velocity pressures at Planes
2 and 5. Since the inlets to the inlet dampers (Planes
1a and 1b) are located shortly downstream of the
traverse planes (Planes 3a and 3b) in an airway of
uniform cross-section, the conditions which exist at
the traverse planes are assumed to exist at the inlets
to the inlet dampers.
Ps1 = Ps3
= (Ps3a + Ps3b)/2
Pv1 is calculated using the total flow rate and the total
area at the inlets to the inlet dampers.
Pv1
= (Q /1096A )2 ρ
1
1
1
8. In order to compare the test results to the quoted
fan curve drawn for operation at 880 rpm and 0.049
lbm/ft3 density, it is necessary to convert the results
to the specified conditions. The basis for calculations
is described in Section 14.
OBSERVATIONS
SITE MEASUREMENTS
pb = 30.12 in. Hg
td3a = 345°F
td3b = 359°F
td5 = 363°F
Ps3a = -18.8 in. wg
Ps3b = -18.3 in. wg
Pv3a = 2.053 in. wg
Pv3b = 2.028 in. wg
Ps5 = -1.6 in. wg
N = 892 rpm
A1a = A1b
= 60.7 ft2
A2 = 115 ft2
A3a = A3b
= 60.7 ft2
A5 = 140 ft2
Blast Area = 80 ft2
The density of the gas, as determined by Orsat
analysis, is 0.0725 lbm/ft3 at 29.92 in. Hg and 70°F.
MEASURED MOTOR DATA
Volts = 4300, 4250, 4200
= 4250 av
Amps = 378, 376, 380
= 378 av
kW
= 2519
MOTOR NAMEPLATE DATA
3000 hp, 3 phase, 60 hertz
4000 volts, 880 rpm, 385 FLA
GENERAL
Inlet box dampers and outlet damper in full open
positions. Fan direct connected to motor. Motor
efficiency data supplied by motor manufacturer.
Pressure loss data supplied by manufacturer of outlet
damper.
CALCULATIONS
DENSITIES
The densities at Planes 3a and 3b are:
⎛ P + 13.6 pb ⎞ ⎛ 70 + 460 ⎞
ρ3a = 0.0725 ⎜ s3a
⎟
⎟⎜
⎝ 13.6 × 29.92 ⎠ ⎝ t d3a + 460 ⎠
⎛ −18.8 + 13.6 × 30.12 ⎞ ⎛ 530 ⎞
= 0.0725 ⎜
⎟ ⎜ 805 ⎟
13.6 × 29.92
⎝
⎠⎝
⎠
3
= 0.0458 lbm/ft
⎛ P + 13.6 pb ⎞ ⎛ 70 + 460 ⎞
ρ3b = 0.0725 ⎜ s3b
⎟
⎟⎜
⎝ 13.6 × 29.92 ⎠ ⎝ t d3b + 460 ⎠
⎛ −18.3 + 13.6 × 30.12 ⎞ ⎛ 530 ⎞
= 0.0725 ⎜
⎟ ⎜ 819 ⎟
13.6 × 29.92
⎝
⎠⎝
⎠
3
= 0.0451 lbm/ft
It is assumed that ρ1a = ρ3a and ρ1b = ρ3b.
The density at Plane 5:
⎛ P + 13.6 pb ⎞ ⎛ 70 + 460 ⎞
ρ5 = 0.0725 ⎜ s5
⎟
⎟⎜
⎝ 13.6 × 29.92 ⎠ ⎝ t d5 + 460 ⎠
⎛ −1.6 + 13.6 × 30.12 ⎞ ⎛ 530 ⎞
= 0.0725 ⎜
⎟ ⎜ 823 ⎟
13.6 × 29.92
⎝
⎠⎝
⎠
= 0.0468 lbm/fft 3
It is assumed that ρ2 = ρ5.
FLOW RATES
V3a = 1096 (Pv3a/ρ3a)0.5
= 1096 (2.053/0.0458)0.5
= 7338 fpm
Q3a = V3aA3a
= 7338 × 60.7
= 445417 cfm
Field Performance Measurement | 193
V3b = 1096 (Pv3b/ρ3b)0.5
= 1096 (2.028/0.0451)0.5
= 7349 fpm
Q3b = V3bA3b
= 7349 × 60.7
= 446084 cfm
Q3 = Q3a + Q3b
= 445417 + 446084
= 891501 cfm
Since the air is divided evenly between the two inlet
boxes:
FAN STATIC PRESSURE
Pv1 = (Q1/1096A1)2 ρ1
= (891501/1096 × 121.4)2 0.0455
= 2.04 in. wg
Pv2 = Pv1 (A1/A2)2 (ρ1/ρ2)
= 2.04 (121.4/115)2 (0.0455/0.0468)
= 2.21 in. wg
Pv5 = Pv1 (A1/A5)2 (ρ1/ρ5)
= 2.04 (121.4/140)2 (0.0455/0.0468)
= 1.49 in. wg
ρ1 = ρ3
= (ρ3a + ρ3b)/2
= (0.0458 + 0.0451)/2
= 0.0455 lbm/ft3
Ps2 + Pv2 = Ps5 + Pv5 + Damper Loss
Q = Q1
= Q3 (ρ3/ρ1)
= 891501 (0.0455/0.0455)
= 891501 cfm
Ps1 = Ps3
= (Ps3a + Ps3b)/2
= (-18.8 - 18.3)/2
= -18.55 in. wg
FAN POWER INPUT
Ps = Ps2 - Ps1 - Pv1
= -1.57 - (-18.55) - 2.04
= 14.94 in. wg
Measured amps/FLA = (378/385)
= 0.98
= 98%
Annex K indicates that Equation A will provide a
reasonably accurate estimate of motor power output
for a 3000 hp motor operating at 98% FLA.
Hmo = 3000 (378/385) (4250/4000)
= 3130 hp
The data supplied by the motor manufacturer indicate
motor efficiency of 94% at the measured power input
of 2519 kW. Using this information:
Hmo = (2519 × 0.94)/0.746
= 3174 hp
The more accurate method of estimating the motor
power output is assumed to be the latter. Since the
fan is direct connected to the motor, there is no drive
loss, and:
H = Hmo
= 3174 hp
194 | Field Performance Measurement
Ps2 = Ps5 + Pv5 + Damper Loss - Pv2
= -1.6 + 1.49 + 0.75 - 2.21
= -1.57 in. wg
CONVERSION TO SPECIFIED CONDITIONS
Qc = 891501 (880/892)
= 879508 cfm
Psc = 14.94 (880/892)2 (0.049/0.0455)
= 15.66 in. wg
Hc = 3174 (880/892)3 (0.049/0.0455)
= 3282 hp
EXAMPLE 2F: AXIAL FAN IN A VENTILATION SYSTEM
3
GUIDE VANES
STATIC PRESSURE TAPS
4
5
SEF 1
2-PIECE ELBOW
(TYPICAL)
INNER
CYLINDER
L1
1
SEF 2
L2
2
COMMENTS
1. The unusual duct arrangement in this example
makes it very difficult to obtain accurate pressure
measurements, and this fact should be understood
before testing begins. Also, the use of a diverging
inlet fitting and a converging outlet fitting with this fan
can pose additional problems. Unless the degrees of
divergence and convergence are moderate, as they
are in this example, the fan performance will be
adversely affected.
2. Determine Pv3 by using the root mean square of
the velocity pressure measurements made in a
traverse of Plane 3, located well downstream in a
straight run of duct, such as shown in the diagram.
Determine Ps3 by averaging the static pressure
measurements made in the same traverse.
Procedures for traverses are described in Section
9.4. Ps3 is used in determining the density at the
traverse plane. Measure the area of the traverse
plane, A3.
3. Determine Ps5 by averaging the pressure
measurements at each of four static pressure taps
located near the end of the duct connection at the fan
outlet. Determine Ps4 by using static pressure taps in
the duct connection at the fan inlet. Measure A4 and
A5, the cross-sectional areas of the duct connections
at the static pressure taps.
4. Measure td3 and tw3 in the traverse plane.
Determine pb for the general vicinity of the fan.
Measure td4. These measurements are used in
determining densities at the various planes of
interest.
5. Measure the fan speed, motor amps, volts, and if
possible, watts. Record all pertinent motor nameplate
data, including volts (NPV) and full load amps (FLA).
If the motor power output is to be estimated by using
the phase current method described in Annex K, it is
not necessary to measure motor watts; however, it
may be necessary to disconnect the drive and
measure the no load amps (NLA) if the motor is not
operating at or near its full load point. Motor
performance data, supplied by the motor
manufacturer, are used in the determination of motor
power output for this example.
6. SEF 1 is due to the effect of insufficient length of
duct between the fan inlet and the elbow upstream of
the fan. SEF 2 is due to the effect of insufficient
length of duct between the fan outlet and the elbow
downstream of the fan. In order to calculate the
values of the SEFs, it is necessary to measure the
inlet area and the outlet area of the fan, A1 and A2;
and the lengths of the inlet and outlet duct
connections, L1 and L2.
Field Performance Measurement | 195
7. To calculate the Fan Static Pressure:
460 volts, 1760 rpm, 24.6 FLA
Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2
GENERAL
Where:
Fan direct connected to motor. Motor efficiency data
supplied by motor manufacturer. Fan speed
measurement was not obtained due to the closed
duct arrangements on both sides of the fan. The
measured amps indicate that the motor is operating
very close to the full load condition, so the rpm was
assumed to be the motor nameplate value of 1760.
Pv1 = Pv3 (A3/A1)2 (ρ3/ρ1)
Ps2 and Ps1 are calculated using measured static
pressure values and constant total pressure
considerations.
Ps1 + Pv1 = Ps4 + Pv4
Ps2 + Pv2 = Ps5 + Pv5
Where each velocity pressure is calculated in a
manner similar to the calculation of Pv1, shown
above.
8. In order to compare the test results to the quoted
fan curve drawn for operation at 1750 rpm and 0.075
lbm/ft3 density, it is necessary to convert the results
to the specified conditions. The basis for the
calculations is described in Section 14.
OBSERVATIONS
SITE MEASUREMENTS
pb = 29.76 in. Hg
td3 = 82.8°F
tw3 = 57.2°F
td4 = 80°F
Ps3 = 0.5 in. wg
Pv3 = 0.783 in. wg
Ps4 = -1.1 in. wg
Ps5 = 0.82 in. wg
A1 = A2
= 7.1 ft2
A3 = A5
= 4.91 ft2
A4 = 6.2 ft2
L1 = 3.0 ft
L2 = 3.5 ft
CALCULATIONS
DENSITIES
For Plane 3 conditions of:
td3 = 82.8°F
tw3 = 57.2°F
p3 = pb + (Ps3/13.6)
= 29.76 + (0.5/13.6)
= 29.80 in. Hg
Use Figure N.1 in Annex N to obtain ρ3 = 0.0728
lbm/ft3.
It is assumed ρ2 = ρ5 = ρ3.
The density at Planes 1 and 4:
ρ1 = ρ 4
⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞
= ρ3 ⎜ s4
⎟⎜
⎟
⎝ 13.6 p3 ⎠ ⎝ t d4 + 460 ⎠
⎛ −1.1 + 13.6 × 29.76 ⎞ ⎛ 542.8 ⎞
= 0.0728 ⎜
⎟ ⎜ 540 ⎟
13.6 × 29.80
⎝
⎠⎝
⎠
= 0.0729 lbm/ft 3
FLOW RATES
V3 = 1096 (Pv3/ρ3)0.5
= 1096 (0.783/0.0728)0.5
= 3594 fpm
MEASURED MOTOR DATA
Volts = 460, 461, 459
= 460 av
Amps = 25.0, 25.0, 24.8
= 24.9 av
kW = 18.0
MOTOR NAMEPLATE DATA
20 hp, 3 phase, 60 hertz
196 | Field Performance Measurement
Q3 = V3A3
= 3594 × 4.91
= 17647 cfm
Q = Q1
= Q3 (ρ3/ρ1)
= 17647 (0.0728/0.0729)
= 17623 cfm
FAN POWER INPUT
Diameter of the fan outlet:
The data supplied by the motor manufacturer indicate
motor efficiency of 87.5% at the measured power
input of 18.0 kW. Using this information:
D2 = (4A2/π)0.5
= (4 × 7.1/π)0.5
= 3.01 ft
Hmo = (18.0 × 0.875)/0.746
= 21.1 hp
Figure 8.1 shows that for velocities of 2500 fpm or
less, the 100% effective duct length is 2.5 diameters:
Since the fan is direct connected to the motor, there
is no drive loss, and:
H = Hmo
= 21.1 hp
SYSTEM EFFECT FACTORS
To determine the value of SEF 1, calculate the
velocity at the fan inlet:
V1 = (Q1/A1)
= (17623/7.1)
= 2482 fpm
= 2.5 × 3.01
= 7.53 ft
The length of the outlet duct in % effective duct
length:
= (L2/7.53) 100
= (3.5/7.53) 100
= 46%
From Figure 8.4, for a vaneaxial fan with a 46%
effective duct length between its discharge and a two
piece elbow, System Effect Curve W applies. From
Figure 7.1 for 2485 fpm velocity and curve W, SEF 2
is less than 0.1 in. and is considered negligible.
Diameter of the fan inlet:
SEF 2 = 0.00
D1 = (4A1/π)0.5
= (4 × 7.1/π)0.5
= 3.01 ft
The length of the duct between the elbow and the fan
inlet in terms of the fan inlet diameter:
= (L1/D1)
= (3.0/3.01)
= 1.00
AMCA Publication 201-90, Figure 9.2 indicates that
for a two piece elbow with a length of duct between
the elbow and the fan inlet equal to 1.00 diameter
System Effect Curve S-T applies. For a velocity of
2482 fpm and curve S-T, Figure 7.1 shows SEF 1 =
0.25 in. wg at 0.075 lbm/ft3. At 0.0729 lbm/ft3:
SEF 1 = 0.25 (0.0729/0.075)
= 0.24 in. wg
For SEF 2, AMCA Publication 201-90, Figures 7.1,
8.1, and 8.4 indicate the following calculations:
Q2 = Q3 (ρ3/ρ2)
= 17647 (0.0728/0.0728)
= 17647 cfm
V2 = Q2/A2
= 17647/7.1
= 2485 fpm
FAN STATIC PRESSURE
Pv5 = Pv3 (A3/A5)2 (ρ3/ρ5)
= 0.783 (4.91/4.91)2 (0.0728/0.0728)
= 0.783 in. wg
Pv2 = Pv3 (A3/A2)2 (ρ3/ρ2)
= 0.783 (4.91/7.1)2 (0.0728/0.0728)
= 0.37 in. wg
Ps2 + Pv2 = Ps5 + Pv5
Ps2 = Ps5 + Pv5 - Pv2
= 0.82 + 0.783 - 0.37
= 1.23 in. wg
Pv4 = Pv3 (A3/A4)2 (ρ3/ρ4)
= 0.783 (4.91/6.2)2 (0.0728/0.0729)
= 0.49 in. wg
Pv1 = Pv3 (A3/A1)2 (ρ3/ρ1)
= 0.783 (4.91/7.1)2 (0.0728/0.0729)
= 0.37 in. wg
Ps1 + Pv1 = Ps4 + Pv4
Ps1 = Ps4 + Pv4 - Pv1
= -1.1 + 0.49 - 0.37
= -0.98 in. wg
Field Performance Measurement | 197
Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2
= 1.23 - (-0.98) - 0.37 + 0.24 + 0
= 2.08 in. wg
CONVERSION TO SPECIFIED CONDITIONS
Qc = 17623 (1750/1760)
= 17523 cfm
Psc = 2.08 (1750/1760)2 (0.075/0.0729)
= 2.12 in. wg
Hc = 21.1 (1750/1760)3 (0.075/0.0729)
= 21.3 hp
198 | Field Performance Measurement
EXAMPLE 2G: HIGH PRESSURE CENTRIFUGAL FAN IN A SERIES
3
2b
STATIC PRESSURE TAPS
1b
2a
FAN B
1a
DAMPER
INLET BOX
INLET BOX
FAN A
SIDE VIEW
COMMENTS
1. The two single inlet fans in this example have been
rated by the manufacturer as a two stage assembly.
Although rated as an assembly, sufficient
measurements are made to provide performance
data for each fan. The damper downstream of the
second fan is not included as part of the rated
assembly. In virtually all cases in which an air flow
control damper, such as the one shown in the
diagram, is included in the system, the point of
operation of major interest and for which the fan has
been selected is at the maximum air flow rate. This
example is no exception. Therefore, it is essential
that the damper be fixed in its full open position for
the duration of the test.
2. Determine Pv3 by using the root mean square of
the velocity pressure measurements made in a
traverse of Plane 3. Determine Ps3 by averaging the
static pressure measurements made in the same
traverse. Procedures for traverses are described in
Section 9.4. Ps3 is used in determining the density at
the traverse plane, A3, which is located at the tip of
the Pitot-static tube.
3. Determine the static pressures at Planes 1a, 1b2a, and 2b. As shown in the diagram, these planes
are located shortly downstream of the inlets and
outlets of the fans, which are the planes of interest. In
each case, the conditions which exist at the plane of
measurements are assumed to exist at the
respective plane of interest because of the close
proximity and the fact that the two planes are equal in
area. The static pressure at each plane may be
determined by averaging the static pressure
measurements at each of four static pressure taps, or
by averaging the static pressure measurements
made in a Pitot-static tube traverse of the plane.
However, due to the turbulence existing in the
regions of the outlets of the fans, it is recommended
that static pressure taps be used at Planes 1b-2a and
2b.
4. Measure td3, tw3, td1b, and td2b; td1a is assumed to be
equal to td3. Determine pb for the general vicinity of
the fan. These measurements are used in
determining densities at the planes of interest.
5. Measure the fan speed and the motor amps, volts,
and if possible, watts for each fan. Record all
pertinent motor nameplate data, including volts
(NPV), and full load amps (FLA). If the motor power
outputs are to be estimated by using the phase
current method described in Annex K, it is not
necessary to measure motor watts; however, it may
be necessary to disconnect the drives and measure
the no load amps (NLA) if the motors are not
operating at or near their full load points. In this
example, a watts input measurement is made for
Field Performance Measurement | 199
each motor and motor performance data, supplied by
the motor manufacturer, are used in determining
motor power outputs.
6. The SEF which would normally be attributed to
insufficient length of duct at the outlet of the first
stage fan does not apply in this case because the
fans have been rated as an assembly.
7. To calculate the static pressure for the two stage
assembly:
Second Stage
Volts = 4080, 4040, 4020
= 4047 av
Amps = 44, 44.5, 45
= 44.5 av
kW
= 272
MOTOR NAMEPLATE DATA
Data identical for each stage:
350 hp, 3 phase, 60 hertz
4000 volts, 1790 rpm, 44.5 FLA
Ps = Ps2b - Ps1a - Pv1a
GENERAL
Where:
Pv1a = Pv3 (A3/A1a
)2 (ρ /ρ
3
1a
)
8. In order to compare the test results to the
performance quoted for the two stage assembly for
operation at 1780 rpm and 0.045 lbm/ft3 density, it is
necessary to convert the results to the specified
conditions. The basis for the calculations is described
in Section 14.
OBSERVATIONS
SITE MEASUREMENTS
= 28.64 in. Hg
= 35°F
= 33°F
= td2a
= 95°F
td2b = 147°F
Pv3 = 0.745 in. wg
Ps3 = -150 in. wg
Ps1b = Ps2a
= -79.5 in. wg
Ps2b = 0.5 in. wg
Na = 1790 rpm, first stage fan speed
Nb = 1790 rpm, second stage fan speed
A1a = A2a = A1b = A2b
= 5.6 ft2
A3 = 4.92 ft2
pb
td3
tw3
td1b
MEASURED MOTOR DATA
First Stage
Volts = 4000, 4040, 4080
= 4040 av
Amps = 44.5, 45, 45.5
= 45 av
kW
= 278
200 | Field Performance Measurement
Fans direct connected to motors. Motor efficiency
data supplied by motor manufacturer.
CALCULATIONS
DENSITIES
For Plane 3 conditions of:
td3 = 35°F
tw3 = 33°F
p3 = pb + (Ps3/13.6)
= 28.64 + (-150/13.6)
= 17.61 in. Hg
Use the modified Apjohn equation for partial vapor
pressure and the density equation based on perfect
gas relationships, both of which are described in
Annex M, and the data in Figure N.2 in Annex N to
calculate the density at Plane 3.
pe = 0.1879 in. Hg
p3 (t d3 − t w3 )
2700
17.61(35 − 33)
= 0.1879 −
2700
= 0.1749 in. Hg
pp = pe −
ρ3 =
=
1.3257( p3 − 0.378 pp )
t d3 + 460
1.3257 (17.61 − 0.378 × 0.1749 )
35 + 460
= 0.0470 lbm/ft 3
Any conversion of velocity pressure to static pressure
which may occur between Planes 3 and 1a can be
ignored with no significant effect on the accuracy of
the test results. Therefore:
Ps1a = Ps3
= -150 in. wg
FAN POWER INPUT
Assuming no change in temperature between Planes
3a and 1a:
At the measured power input values of 278 kW and
272 kW, the data supplied by the motor manufacturer
indicate efficiency of 95% for each motor.
ρ1a = ρ3
= 0.0470 lbm/ft3
Hmoa = (278 × 0.95)/0.746
= 354 hp
To provide information regarding the flow rates
between stages and leaving the second stage,
additional density values are calculated as follows:
Hmob = (272 × 0.95)/0.746
= 346 hp
ρ1b = ρ2a
⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞
= ρ3 ⎜ s1b
⎟⎜
⎟
⎝ 13.6 p3
⎠ ⎝ t d4 + 460 ⎠
⎛ −79.5 + 13.6 × 28.64 ⎞ ⎛ 495 ⎞
= 0.0470 ⎜
⎟ ⎜ 555 ⎟
13.6 × 17.61
⎝
⎠⎝
⎠
Since each fan is direct connected to its motor, there
are no drive losses and:
Ha
= Hmoa
= 354 hp
Hb
= Hmob
= 346 hp
= 0.0543 lbm/fft 3
FAN STATIC PRESSURE
⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞
ρ2b = ρ3 ⎜ s2b
⎟⎜
⎟
⎝ 13.6 p3
⎠ ⎝ t d2b + 460 ⎠
⎛ 0.5 + 13.6 × 28.64 ⎞ ⎛ 495 ⎞
= 0.0470 ⎜
⎟⎜
⎟
⎝ 13.6 × 17.61 ⎠ ⎝ 607 ⎠
= 0.0624 lbm/ftt 3
FLOW RATES
V3 = 1096 (Pv3/ρ3)0.5
= 1096 (0.745/0.0470)0.5
= 4364 fpm
Q3 = V3A3
= 4364 × 4.92
= 21471 cfm
Q
= Q1a
= Q3 (ρ3/ρ1a)
= 21471 (0.0470/0.0470)
= 21471 cfm
Q1b = Q2a
= Q3 (ρ3/ρ2a)
= 21471 (0.0470/0.0543)
= 18584 cfm
Pv1a = Pv3 (A3/A1a)2 (ρ3/ρ1a)
= 0.745 (4.92/5.6)2 (0.0470/0.0470)
= 0.575 in. wg
The static pressure for the two stage assembly:
Ps
= Ps2b - Ps1a - Pv1a
= 0.5 - (-150) - 0.575
= 149.9 in. wg
CONVERSION TO SPECIFIED CONDITIONS
Qc
= 21471 (1780/1790)
= 21351 cfm
Psc
= 149.9 (1780/1790)2 (0.045/0.0470)
= 141.9 in. wg
Hac = 354 (1780/1790)3 (0.045/0.0470)
= 333 hp
Hbc = 346 (1780/1790)3 (0.045/0.0470)
= 326 hp
Q2b = Q3 (ρ3/ρ2b)
= 21471 (0.0470/0.0624)
= 16172 cfm
Field Performance Measurement | 201
EXAMPLE 3A: CENTRIFUGAL FAN IN AN EXHAUST SYSTEM
AIR INTAKE VENTS
BACKDRAFT DAMPER
SEF 1
3a
2
3c
3b
1
STATIC PRESSURE TAPS
PLAN VIEW
COMMENTS
1. This fan, as supplied and rated by the
manufacturer, does not include the backdraft damper.
2. Normally, velocity pressure measurements would
be made in a single plane, located in a duct common
to all branches. In this example, a measurement
plane which provides a satisfactory velocity profile
cannot be located within the short length of duct
between the point of connection of the branch ducts
and the fan inlet. The alternative, as indicated in the
diagram, is to make a velocity pressure
measurement traverse in the longest available duct
run of each branch. The velocity pressure for each
branch is determined by using the root mean square
of the velocity pressure measurements made in the
traverse. The static pressure at each traverse plane
is determined by averaging the static pressure
measurements made in the same traverse. These
static pressure values are used in determining the
densities at the traverse planes. Procedures for
traverses are described in Section 9.4. In order to
determine the air flow rates it is necessary to
measure the area of each traverse point.
3. Ps1, the static pressure at the fan inlet may be
determined by averaging the static pressure
measurements at each of four static pressure taps or
by averaging the static pressure measurements
made in a Pitot-static tube traverse of Plane 1. If a
202 | Field Performance Measurement
Pitot-static tube is used, it should be positioned well
within the inlet collar in which Plane 1 is located.
Measure the area of Plane 1 for use in calculating
Pv1. The static pressure at the outlet of the backdraft
damper is zero gauge pressure, referred to the
atmospheric pressure in the region of the outlet of the
backdraft damper. In situations such as this example,
the air may be discharging from the damper into a
region in which the atmospheric pressure is
somewhat different from that to which all other
pressure measurements are referred. When this
possibility exists, it is essential that the static
pressure in the region of the discharging air be
measured, referred to the same atmospheric
pressure as used in all other pressure
measurements.
4. Measure the dry-bulb and wet-bulb temperatures
at each velocity traverse plane and the dry-bulb
temperature at Plane 1. In this example, td2 is
assumed to be equal to td1. Determine pb for the
general vicinity of the fan. These measurements are
used in determining densities at the planes of
interest.
5. Measure the fan speed and the motor amps, volts,
and if possible, watts. Record all pertinent motor
nameplate data, including volts (NPV) and full load
amps (FLA). If the motor power output is to be
estimated by using the phase current method
described in Annex K, it is not necessary to measure
motor watts; however, it may be necessary to
disconnect the drive and measure the no load amps
(NLA) if the motor is not operating at or near its full
load point. Refer to Annex K.
6. SEF 1 is due to the effect of there being no duct
at the fan outlet. In order to calculate the value of
SEF 1, it is necessary to measure the outlet area of
the fan, A2, and the blast area of the fan.
7. Determine the backdraft damper pressure loss by
using the performance ratings supplied by the
manufacturer and the pressure loss multiplier data in
Figure 8.7 of AMCA Publication 201-90. The use of
the multiplier is indicated because the damper is
mounted directly to the fan outlet.
Pv3a = 0.765 in. wg
Pv3b = 0.88 in. wg
Pv3c = 0.86 in. wg
N = 800 rpm
A1 = 16.8 ft2
A2 = 13.8 ft2
A3a = 5.4 ft2
A3b = A3c
= 3.0 ft2
Blast Area = 11.0 ft2
MEASURED MOTOR
Volts = 460, 458, 462
= 460 av
Amps = 28, 27, 26
= 27 av
NLA = 14.7
8. To calculate the Fan Static Pressure:
MOTOR NAMEPLATE DATA
Ps = Ps2 - Ps1 - Pv1 + SEF 1
Where:
25 hp, 3 phase, 60 hertz
460 volts, 1760 rpm, 32 FLA
Pv1 = (Q1/1096 A1)2 ρ1
GENERAL
Q1 = Q3a (ρ3a/ρ1) + Q3b (ρ3b/ρ1) + Q3c (ρ3c/ρ1)
Fan connected to motor through belt drive. Pressure
loss data supplied by manufacturer of backdraft
damper.
Ps2 is the sum of the static pressure in the region of
the damper outlet, which was measured as zero, and
the backdraft damper pressure loss.
9. In order to compare the test results to the quoted
fan curve drawn for operation at 810 rpm and 0.075
lbm/ft3 density, it is necessary to convert the results
to the specified conditions. The basis for the
calculations is described in Section 14.
OBSERVATIONS
SITE MEASUREMENTS
pb = 29.8 in. Hg
td1 = 72°F
tw1 = 62°F
td3a = 77°F
tw3a = 67°F
td3b = 65°F
tw3b = 56°F
td3c = 70°F
tw3c = 62°F
Ps1 = -1.00 in. wg
Ps3a = -0.80 in. wg
Ps3b = -0.45 in. wg
Ps3c = -0.040 in. wg
CALCULATIONS
DENSITIES
Since the static pressure values at Planes 1, 3a, 3b,
and 3c are very small, no appreciable error will occur
by using the barometric pressure instead of the
absolute pressure at each plane in the determination
of the densities. The densities at these planes are
obtained by using Figure N.1 in Annex N.
ρ1 = 0.0739 lbm/ft3
ρ3a = 0.0731 lbm/ft3
ρ3b = 0.0750 lbm/ft3
ρ3c = 0.0741 lbm/ft3
FLOW RATES
V3a = 1096 (Pv3a/ρ3a)0.5
= 1096 (0.765/0.0731)0.5
= 3546 fpm
V3b = 1096 (Pv3b/ρ3b)0.5
= 1096 (0.88/0.0750)0.5
= 3754 fpm
Field Performance Measurement | 203
V3c = 1096 (Pv3c/ρ3c)0.5
= 1096 (0.86/0.0741)0.5
= 3734 fpm
Q3a = V3aA3a
= 3546 × 5.4
= 19148 cfm
Q3b = V3bA3b
= 3754 × 3.0
= 11262 cfm
Q3c = V3cA3c
= 3734 × 3.0
= 11202 cfm
Q = Q1
= Q3a ( ρ3a / ρ1 ) + Q3 b ( ρ3 b / ρ1 ) + Q3c ( ρ3c / ρ1 )
⎛ 0.0731 ⎞
⎛ 0.0750 ⎞
⎛ 0.0741 ⎞
= 19148 ⎜
⎟ + 11262 ⎜ 0.0739 ⎟ + 11202 ⎜ 0.0739 ⎟
0
0739
.
⎝
⎠
⎝
⎠
⎝
⎠
= 41603 cfm
SYSTEM EFFECT FACTOR
AMCA Publication 201-90, Figures 7.1 and 8.3
indicate the following calculations:
Q2 = Q1
= 41603 cfm
It is assumed that ρ2 = ρ1.
V2 = (Q2/A2)
= (41603/13.8)
= 3015 fpm
Blast area ratio = Blast area/A2
= 11.0/13.8
= 0.80
For a blast area ratio of 0.8 and no duct, Figure 8.3
shows System Effect Curve T-U applies. For 3015
fpm velocity and curve T-U, Figure 7.1 shows SEF 1
= 0.27 in. wg at 0.075 lbm/ft3 density. At 0.0739
lbm/ft3:
FAN POWER INPUT
SEF 1 = 0.27 (0.0739/0.075)
= 0.27 in. wg
Measured amps/FLA = (27/32)
= 0.84
= 84%
BACKDRAFT DAMPER LOSS MULTIPLIER
Annex K indicates that the average of the results of
Equation A and Equation B will provide a reasonably
accurate estimate of motor power output for a 25 hp
motor operating at 84% FLA.
Eqn A = 25 (27/32) (460/460)
= 21.1 hp
Eqn B = 25 [(27 - 14.7)/(32 - 14.7)] (460/460)
= 17.8 hp
Hmo
= (21.1 + 17.8)/2
= 19.45 hp
The data supplied by the manufacturer of the damper
indicate that the pressure loss for the damper, ∆Ps, is
0.4 in. wg at the flow rate of 41603 cfm at 0.075
lbm/ft3 density. AMCA Publication 201-90, Figure 8.7
indicates a ∆Ps multiplier of 1.9 for a damper which is
mounted directly to the outlet of a fan which has a
blast area ratio of 0.8.
Backdraft damper loss = ∆Ps × 1.9 × (ρ2/0.075)
= 0.4 × 1.9 (0.0739/0.075)
= 0.75 in. wg
FAN STATIC PRESSURE
Figure L.1 in Annex L indicates estimated belt drive
loss of 4.8%.
Pv1 = (Q1/1096 A1)2 ρ1
= [41603/(1096 × 16.8)]2 0.0739
= 0.38 in. wg
HL = 0.048 Hmo
= 0.048 × 19.45
= 0.93 hp
Ps2 is equal to the static pressure at the outlet of the
damper, which is zero, plus the damper loss.
H = Hmo - HL
= 19.45 - 0.93
= 18.52 hp
204 | Field Performance Measurement
Ps2 = 0 + damper loss
= 0 + 0.75
= 0.75 in. wg
Ps = Ps2 - Ps1 - Pv1 + SEF 1
= 0.75 - (-1.0) - 0.38 + 0.27
= 1.64 in. wg
CONVERSION TO SPECIFIED CONDITIONS
Qc = 41603 (810/800)
= 42123 cfm
Psc = 1.64 (810/800)2 (0.075/0.0739)
= 1.71 in. wg
Hc = 18.52 (810/800)3 (0.075/0.0739)
= 19.51 hp
Field Performance Measurement | 205
EXAMPLE 3B: AXIAL FAN IN AN EXHAUST SYSTEM
3
2-PIECE ELBOW
SEF 1
L1
1
STATIC
PRESSURE TAPS
GUIDE VANES
INNER CYLINDER
2
L2
SEF 2
5
PLAN VIEW
COMMENTS
1. Determine Pv3 by using the root mean square of
the velocity pressure measurements made in a
traverse of Plane 3, located near the end of a straight
run of duct, as shown in the diagram. Determine Ps3
by averaging the static pressure measurements
made in the same traverse. Procedures for traverses
are described in Section 9.4. Ps3 is used in
determining the density at the traverse plane.
Measure the area of the traverse plane, A3, which is
located at the tip of the Pitot-static tube.
2. Determine Ps5 by averaging the pressure
measurements at each of four static pressure taps
located near the end of the duct connection at the fan
outlet. Determine Ps1 by using a Pitot-static tube or
static pressure taps in the duct connection at the fan
inlet. If a Pitot-static tube is used, it should not project
into the upstream elbow but be located well within the
length of the duct connection.
3. Measure td3 and tw3 in the traverse plane; td1 is
assumed to be equal to td3. Determine pb for the
general vicinity of the fan. Measure td5. These
measurements are used in determining densities at
the planes of interest.
4. Measure the fan speed and the motors amps,
volts, and if possible, watts. Record all pertinent
206 | Field Performance Measurement
motor nameplate data, including volts (NPV), and full
load amps (FLA). If the motor power output is to be
estimated by using the phase current method
described in Annex K, it is not necessary to measure
motor watts; however, it may be necessary to
disconnect the drive and measure the no load amps
(NLA) if the motor is not operating at or near its full
load point. Refer to Annex K.
5. SEF 1 is due to the effect of insufficient length of
duct between the fan inlet and the elbow upstream of
the fan. SEF 2 is due to the effect of insufficient
length of duct between the fan outlet and the elbow
downstream of the fan. In order to calculate the
values of the SEFs, it is necessary to measure the
inlet area and the outlet area of the fan, A1 and A2;
and the lengths of the inlet and outlet duct
connections, L1 and L2.
6. To calculate the Fan Static Pressure:
Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2
Where: Pv1 = Pv3
Since:
A1 = A3
And:
CALCULATIONS
ρ1 = ρ3
DENSITIES
Due to the close proximity of Planes 2 and 5 and the
fact that there is no change in area between the two
planes, all conditions which exist at Plane 5 are
assumed to exist at Plane 2.
For Plane 3 conditions of:
Therefore:
td3 = 72°F
tw3 = 66°F
p3 = pb + (Ps3/13.6)
= 29.20 + (-1.92/13.6)
= 29.06 in. Hg
Ps2 = Ps5
7. In order to compare the test results to the quoted
fan curve drawn for operation at 1730 rpm and 0.075
lbm/ft3 density, it is necessary to convert the results
to the specified conditions. The basis for the
calculations is described in Section 14.
OBSERVATIONS
SITE MEASUREMENTS
pb = 29.20 in. Hg
td3 = 72°F
tw3 = 66°F
td5 = 73°F
Ps1 = -2.02 in. wg
Ps3 = -1.92 in. wg
Pv3 = 0.35 in. wg
Ps5 = 0.10 in. wg
N = 1710 rpm
A1 = A2 = A3 = A5
= 2.64 ft2
L1 = 1.5 ft, length of inlet duct
L2 = 2.25 ft, length of the outlet duct
Use Figure N.1 in Annex N to obtain ρ3 = 0.0719
lbm/ft3.
Assume that td1 = td3.
⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞
ρ1 = ρ3 ⎜ s1
⎟⎜
⎟
⎝ 13.6 p3 ⎠ ⎝ t d1 + 460 ⎠
⎛ −2.02 + 13.6 × 29.20 ⎞ ⎛ 532 ⎞
= 0.0719 ⎜
⎟ ⎜ 532 ⎟
13.6 × 29.06
⎝
⎠⎝
⎠
3
= 0.0719 lbm/ft
Assume that td2 = td5 and Ps2 = Ps5.
ρ 2 = ρ5
⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞
= ρ3 ⎜ s5
⎟⎜
⎟
⎝ 13.6 p3 ⎠ ⎝ t d5 + 460 ⎠
⎛ 0.10 + 13.6 × 29.20 ⎞ ⎛ 532 ⎞
= 0.0719 ⎜
⎟ ⎜ 533 ⎟
13.6 × 29.06
⎝
⎠⎝
⎠
= 0.0721 lbm/ft 3
MEASURED MOTOR DATA
FLOW RATES
Volts = 227, 229, 228
= 228 av
Amps = 12.2, 12.3, 12.4
= 12.3 av
NLA = 7
V3 = 1096 (Pv3/ρ3)0.5
= 1096 (0.35/0.0719)0.5
= 2418 fpm
MOTOR NAMEPLATE DATA
5 hp, 3 phase, 60 hertz
230 volts, 1760 rpm, 14.0 FLA
GENERAL
Fan connected to motor through belt drive.
Q3 = V3A3
= 2418 × 2.64
= 6384 cfm
Q = Q1
= Q3 (ρ3/ρ1)
= 6384 (0.0719/0.0719)
= 6384 cfm
Q2 = Q5
= Q3 (ρ3/ρ5)
= 6384 (0.0719/0.0721)
= 6366 cfm
Field Performance Measurement | 207
FAN POWER INPUT
Measured amps/FLA = (12.3/14.0)
= 0.88
= 88%
Annex K indicates that the average of the results of
Equation A and Equation B will provide a reasonably
accurate estimate of motor power output for a 5 hp
motor operating at 88% FLA.
R-S, Figure 7.1 shows SEF 1 = 0.24 in. wg at 0.075
lbm/ft3 density. At 0.0719 lbm/ft3:
SEF 1 = 0.24 (0.0719/0.075)
= 0.23 in. wg
For SEF 2, AMCA Publication 201-90, Figures 7.1,
8.1, and 8.4 indicate the following calculations:
Eqn A = 5 (12.3/14) (228/230)
= 4.35 hp
V2 = (Q2/A2)
= (6366/2.64)
= 2411 fpm
Eqn B = 5 [(12.3 - 7)/(14 - 7)] (228/230)
= 3.75 hp
The diameter of the fan outlet:
Hmo
= (4.35 + 3.75)/2
= 4.05 hp
Figure L.1 in Annex L indicates estimated belt drive
loss of 6.3%.
HL = 0.063 Hmo
= 0.063 × 4.05
= 0.26 hp
H = Hmo - HL
= 4.05 - 0.26
= 3.79 hp
SYSTEM EFFECT FACTORS
To determine the value of SEF 1, calculate the
velocity at the fan inlet:
V1 = (Q1/A1)
= (6384/2.64)
= 2418 fpm
D2 = (4A2/π)0.5
= (4 × 2.64/π)0.5
= 1.83 ft
Figure 8.1 shows that for velocities of 2500 fpm or
less, the 100% effective duct length is 2.5 diameters:
= 2.5 × 1.83
= 4.58 ft
The length of the outlet duct in % effective duct
length:
= (L2/4.58) 100
= (2.25/4.58) 100
= 49%
From Figure 8.4, for a vaneaxial fan with a 49%
effective duct length between its discharge and a two
piece elbow, System Effect Curve W applies. From
Figure 7.1, for 2411 fpm velocity and curve W, SEF 2
is less than 0.1 in. wg, and is considered negligible.
SEF 2 = 0.00
Calculate the diameter of the fan inlet:
FAN STATIC PRESSURE
D1 = (4A1/π)0.5
= (4 × 2.64/π)0.5
= 1.83 ft.
Calculate the length of duct between the elbow and
the fan inlet in terms of the fan inlet diameter:
= (L1/D1)
= (1.5/1.83)
= 0.82
AMCA Publication 201-90, Figure 9.2, indicates that
for a vaneaxial fan with a two piece elbow with a
length of duct between the elbow and the fan inlet
equal to 0.8 diameters, System Effect Curve R-S
(estimated) applies. For 2418 fpm velocity and curve
208 | Field Performance Measurement
Since:
A1 = A3
ρ1 = ρ3
Pv1 = Pv3
Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2
= 0.10 - (-2.02) - 0.35 + 0.23 + 0.00
= 2.00 in. wg
CONVERSION TO SPECIFIED CONDITIONS
Qc = 6384 (1730/1710)
= 6459 cfm
Psc = 2.00 (1730/1710)2 (0.075/0.0719)
= 2.14 in. wg
Hc = 3.79 (1730/1710)3 (0.075/0.0719)
= 4.09 hp
Field Performance Measurement | 209
EXAMPLE 3C: CENTRIFUGAL FAN IN A SCRUBBER SYSTEM
3
WET CELL SCRUBBER
1
2
SEF 1
PLAN VIEW
SIDE VIEW
COMMENTS
1. Determine Pv3 by using the root mean square of
the velocity pressure measurements made in a
traverse of Plane 3, located in the duct connection at
the fan inlet, as shown in the diagram. Determine Ps3
by averaging the static pressure measurements
made in the same traverse. Procedures for traverses
are described in Section 9.4. Ps3 is used in
determining the density at the traverse plane.
Measure the area of the traverse plane, A3, which is
located at the tip of the Pitot-static tube. In locating
Plane 3 downstream of the scrubber, changes in the
composition of the air as a result of the action of the
scrubber are properly taken into account in the
determination of fan air flow rate. Due to the close
proximity of Planes 1 and 3, and the fact that there is
no change in area between the two planes, the
conditions which exist at Plane 3 are assumed to
exist at Plane 1.
amps (FLA). If the motor power output is to be
estimated by using the phase current method
described in Annex K, it is not necessary to measure
motor watts; however, it may be necessary to
disconnect the drive and measure the no load amps
(NLA) if the motor is not operating at or near its full
load point. Refer to Annex K.
2. Ps2, the static pressure at the fan outlet, is zero.
Pv1 = Pv3
Ps1 = Ps3
Ps2 = 0
3. Measure td3 and tw3 in the traverse plane.
Determine pb for the general vicinity of the fan.
Measure td2. These measurements are used in
determining densities at the planes of interest.
4. Measure the fan speed and the motor amps, volts,
and if possible, watts. Record all pertinent motor
nameplate data, including volts (NPV), and full load
210 | Field Performance Measurement
5. SEF 1 is due to the effect of there being no duct
at the fan outlet. In order to calculate the value of
SEF 1, it is necessary to measure the outlet area of
the fan, A2, and the blast area of the fan.
6. To calculate the Fan Static Pressure:
Ps = Ps2 - Ps1 - Pv1 + SEF 1
Where:
7. In order to compare the test results to the quoted
fan curve drawn for operation at 1700 rpm and 0.071
lbm/ft3 density, it is necessary to convert the results
to the specified conditions. The basis for the
calculations is described in Section 14.
OBSERVATIONS
SITE MEASUREMENTS
pb = 29.80 in. Hg
td3 = 65°F
tw3 = 64°F
td2 = 70°F
Ps3 = -8.0 in. wg
Pv3 = 0.337 in. wg
N = 1672 rpm
A1 = A3
= 7.06 ft2
A2 = 5.15 ft2
Blast Area = 3.67 ft2
MEASURED MOTOR DATA
Volts = 450, 458, 462
= 457 av
Amps = 44, 45, 44.5
= 44.5 av
MOTOR NAMEPLATE DATA
40 hp, 3 phase, 60 hertz
460 volts, 1780 rpm, 49 FLA
GENERAL
Fan connected to motor through belt drive.
CALCULATIONS
⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞
ρ2 = ρ3 ⎜ s2
⎟⎜
⎟
⎝ 13.6 p3 ⎠ ⎝ t d2 + 460 ⎠
⎛ 0 + 13.6 × 29.80 ⎞ ⎛ 525 ⎞
= 0.0732 ⎜
⎟⎜
⎟
⎝ 13.6 × 29.21 ⎠ ⎝ 530 ⎠
= 0.0740 lbm/ft 3
FLOW RATES
V3 = 1096 (Pv3/ρ3)0.5
= 1096 (0.337/0.0732)0.5
= 2352 fpm
Q3 = V3A3
= 2353 × 7.06
= 16605 cfm
Q = Q1
= Q3 (ρ3/ρ1)
= 16605 (0.0732/0.0732)
= 16605 cfm
Q2 = Q3 (ρ3/ρ2)
= 16605 (0.0732/0.0740)
= 16425 cfm
FAN POWER INPUT
Measured amps/FLA = (44.5/49)
= 0.91
= 91%
DENSITIES
Annex K indicates that Equation A will provide a
reasonably accurate estimate of motor power output
for a 40 hp motor operating at 91% FLA.
For Plane 3 conditions of:
td3 = 65°F
tw3 = 64°F
Hmo = 40 (44.5/49) (457/460)
= 36.1 hp
p3 = pb + (Ps3/13.6)
= 29.80 + (-8.0/13.6)
= 29.21 in. Hg
Use Figure N.1 in Annex N to obtain ρ3 = 0.0732
lbm/ft3.
It is assumed that:
td1 = td3
Ps1 = Ps3
ρ1 = ρ3
Figure L.1 in Annex L indicates estimate belt drive
loss of 4.5%.
HL = 0.045 Hmo
= 0.045 × 36.1
= 1.6 hp
H = Hmo - HL
= 36.1 - 1.6
= 34.5 hp
SYSTEM EFFECT FACTOR
AMCA Publication 201-90, Figures 7.1 and 8.3,
indicate the following calculations:
Field Performance Measurement | 211
V2 = (Q2/A2)
= (16425/5.15)
= 3189 fpm
Blast area ratio = Blast area/A2
= 3.67/5.15
= 0.71
For a blast area ratio of 0.7 and no duct, Figure 8.3
shows System Effect Curve S applies. For 3189 fpm
velocity and curve S, Figure 7.1 shows SEF 1 = 0.5
in. wg at 0.075 lbm/ft3 density. At 0.0740 lbm/ft3:
SEF 1 = 0.5 (0.074/0.075)
= 0.49 in. wg
FAN STATIC PRESSURE
Pv1 = Pv3
= 0.337 in. wg
Ps = Ps2 - Ps1 - Pv1 + SEF 1
= 0 - (-8.0) - 0.337 + 0.49
= 8.15 in. wg
CONVERSION TO SPECIFIED CONDITIONS
Qc = 16605 (1700/1672)
= 16883 cfm
Psc = 8.15 (1700/1672)2 (0.071/0.0732)
= 8.17 in. wg
Hc = 34.5 (1700/1672)3 (0.071/0.0732)
= 35.2 hp
212 | Field Performance Measurement
EXAMPLE 3D: CENTRIFUGAL ROOF VENTILATOR WITH DUCTED INLET
2
1
BACKDRAFT DAMPER
4
STATIC PRESSURE TAPS
3a
SIDE VIEW
3b
COMMENTS
1. This centrifugal roof ventilator, as supplied and
rated by the manufacturer, does not include the
backdraft damper. It is essential that the backdraft
damper blades be fixed in their full open positions,
otherwise uneven velocity distribution will occur at
the inlet to the ventilator, adversely affecting its
performance.
2. Normally, velocity pressure measurements would
be made in a single plane, located in a duct common
to all branches. In this example, a measurement
plane which provides a satisfactory velocity profile
cannot be located within the short length of duct
between the point of connection of the branch ducts
and the ventilator inlet. The alternative, as indicated
in the diagram, is to make a velocity pressure
measurement traverse in each branch. The velocity
pressure for each branch is determined by using the
root mean square of the velocity pressure
measurements made in the traverse. The static
pressure at each traverse plane is determined by
averaging the static pressure measurements made in
the same traverse. These static pressure values are
used in determining the densities at the traverse
planes. Procedures for traverses are described in
Section 9.4. In order to determine the air flow rates, it
is necessary to measure the area of each traverse
plane.
3. Ps4 may be determined by averaging the static
pressure measurements at each of four static
pressure taps or by averaging the static pressure
measurements made in a Pitot-static tube traverse of
Plane 4. If a Pitot-static tube is used, it should be
positioned well within the duct in which Plane 4 is
located, and not project into the upstream elbows.
Measure the area of Plane 1 for use in calculating
Pv1. In this example, A4 = A1. Ps2, the static pressure
at the outlet of the ventilator, is zero gauge pressure,
referred to the atmospheric pressure in the region of
the ventilator outlet. In situations such as this
example, the air may be discharging from the
ventilator into a region in which the atmospheric
pressure is somewhat different from that to which all
other pressure measurements are referred. When
this possibility exists, it is essential that the static
pressure in the region of the discharging air be
measured, referred to the same atmospheric
pressure as used in all other pressure
measurements. In this case, Ps2 was measured as
zero.
4. Measure the dry-bulb and wet-bulb temperatures
at each velocity traverse plane. In this example, td1
and td4 are assumed to be equal to td3a. Determine pb
for the general vicinity of the fan. These
measurements are used in determining densities at
the planes of interest.
Field Performance Measurement | 213
5. Measure the fan speed and the motor amps, volts,
and if possible, watts. Record all pertinent motor
nameplate data, including volts (NPV) and full load
amps (FLA). If the motor power output is to be
estimated by using the phase current method
described in Annex K, it is not necessary to measure
motor watts; however, it may be necessary to
disconnect the drive and measure the no load amps
(NLA) if the motor is not operating at or near its full
load point. Refer to Annex K.
MEASURED MOTOR DATA
Volts = 450, 455, 460
= 455 av
Amps = 5.7, 5.85, 5.9
= 5.82 av
MOTOR NAMEPLATE DATA
5 hp, 3 phase, 60 hertz
460 volts, 1780 rpm, 5.95 FLA
6. Determine the backdraft damper pressure loss by
using the performance ratings supplied by the
manufacturer.
GENERAL
7. To calculate the Fan Static Pressure:
Fan connected to motor through belt drive. Pressure
loss data supplied by manufacturer of backdraft
damper.
Ps = Ps2 - Ps1 - Pv1
CALCULATIONS
Where:
DENSITIES
Pv1
= (Q /1096 A )2 ρ
1
1
1
Q1 = Q3a (ρ3a/ρ1) + Q3b (ρ3b/ρ1)
For Planes 3a and 3b conditions of:
Ps1 = Ps4 - backdraft damper pressure loss
Ps2 = 0
td3a = td3b
= 72°F
tw3a = tw3b
= 66°F
8. In order to compare the test results to the quoted
fan curve drawn for operation at 620 rpm and 0.075
lbm/ft3 density, it is necessary to convert the results
to the specified conditions. The basis for the
calculations is described in Section 14.
p3a = p3b
= pb + (Ps3a/13.6)
= 29.20 + (-0.85/13.6)
= 29.14 in. Hg
OBSERVATIONS
SITE MEASUREMENTS
pb = 29.20 in. Hg
td3a = td3b
= 72°F
tw3a = tw3b
= 66°F
Ps2 = 0 in. wg
Ps4 = -0.88 in. wg
Ps3a = Ps3b
= -0.85 in. wg
Pv3a = 0.27 in. wg
Pv3b = 0.275 in. wg
N = 625 rpm
A1 = A4
= 7.9 ft2
A3a = 3.4 ft2
A3b = 3.3 ft2
214 | Field Performance Measurement
Use Figure N.1 in Annex N to obtain:
ρ3a = ρ3b
= 0.0721 lbm/ft3
It is assumed that:
td1 = td4 = td3a = td3b
Since the differences in the static pressures at
Planes 1, 3a, and 4 are very small, no appreciable
error will occur by assuming:
ρ1 = ρ4 = ρ3a = ρ3b
FLOW RATES
V3a = 1096 (Pv3a/ρ3a)0.5
= 1096 (0.27/0.0721)0.5
= 2121 fpm
V3b = 1096 (Pv3b/ρ3b)0.5
= 1096 (0.275/0.0721)0.5
= 2140 fpm
Q3a = V3aA3a
= 2121 × 3.4
= 7211 cfm
Q3b = V3bA3b
= 2140 × 3.3
= 7062 cfm
BACKDRAFT DAMPER LOSS
The data supplied by the manufacturer of the damper
indicate that the pressure loss for the damper, ∆Ps, is
0.22 in. wg at the flow rate of 14273 cfm at 0.075
lbm/ft3 density.
Backdraft damper loss = ∆Ps (ρ4/0.075)
= 0.22 (0.0721/0.075)
= 0.21 in. wg
FAN STATIC PRESSURE
Q
= Q1
= Q3a (ρ3a/ρ1) + Q3b (ρ3b/ρ1)
= 7211 (0.0721/0.0721) + 7062 (0.0721/0.0721)
= 14273 cfm
FAN POWER INPUT
Measured amps/FLA = (5.82/5.95)
= 0.98
= 98%
Annex K indicates that Equation A will provide a
reasonably accurate estimate of motor power output
for a 5 hp motor operating at 98% FLA.
Pv1 = (Q1/1096 A1)2 ρ1
= [14273/(1096 × 7.9)]2 0.0721
= 0.20 in. wg
Ps1 = Ps4 - damper loss
= -0.88 - 0.21
= -1.09 in. wg
Ps = Ps2 - Ps1 - Pv1
= 0 - (-1.09) - 0.20
= 0.89 in. wg
CONVERSION TO SPECIFIED CONDITIONS
Hmo = 5 (5.82/5.95) (455/460)
= 4.84 hp
Qc = 14273 (620/625)
= 14159 cfm
Figure L.1 in Annex L indicates estimated belt drive
loss of 5.8%.
Psc = 0.89 (620/625)2 (0.075/0.0721)
= 0.91 in. wg
HL = 0.058 Hmo
= 0.058 × 4.84
= 0.28 hp
Hc = 4.56 (620/625)3 (0.075/0.0721)
= 4.63 hp
H
= Hmo - HL
= 4.84 - 0.28
= 4.56 hp
Field Performance Measurement | 215
EXAMPLE 4A: CENTRIFUGAL FAN IN A BUILT-UP AIR CONDITIONING UNIT
2
4
RETURN
AIR
SPRAY
SECTION
5
SEF 2
PLAN VIEW
3a
FAN SECTION
+
+
OUTSIDE
AIR
SEF 1
L
+
+
+
+
+
+
+
+
PREHEAT COILS
FILTER SECTION
DIFFUSER
PLATE
REHEAT COIL
3b
SIDE VIEW
COMMENTS
1. This is an air conditioning unit which has been
assembled at the installation site. The subject of the
test is the fan, which is rated by the manufacturer as
free-standing, unencumbered by the cabinet in which
it has been installed. The fan performance ratings are
based on operation with the fan outlet ducted. Before
proceeding with the test, it is essential that all
dampers--outside air, return air, mixing box,
multizone, face and bypass or volume control--be
fixed in the positions agreed upon by all interested
parties as being applicable for the installation. Also,
the temperatures of the heating coils must be kept
constant throughout the test period. It may be
necessary to lock out, disconnect, or otherwise
modify automatic control devices in order to prevent
the positions of the dampers and temperatures of the
coils from changing during the test. Refer to Section
17.4.3 for additional considerations affecting the test
procedure for fans in this type of installation.
2. Normally, velocity pressure measurements would
be made in a single plane, located in a duct common
to all branches. In this example, a measurement
plane which provides a satisfactory velocity profile
cannot be located upstream of the fan or between the
point of connection of the branch ducts and the fan
outlet. The alternative, as indicated in the diagram, is
to make a velocity pressure measurement traverse in
each branch. The velocity pressure for each branch
216 | Field Performance Measurement
is determined by using the root mean square of the
velocity pressure measurements made in the
traverse. the static pressure at each traverse plane is
determined by averaging the static pressure
measurements made in the same traverse. These
static pressure values are used in determining the
densities at the traverse planes. Procedures for
traverses are described in Section 9.4. In order to
determine the air flow rates, it is necessary to
measure the area of each traverse plane.
3. Determine Ps4 by averaging the static pressure
measurements made in a traverse of Plane 4.
Determine Ps5 in a similar manner. Pitot-static tube
traverses are used in determining these static
pressures because the installation of suitable
pressure taps is usually prevented by the insulating
material encountered in this type of equipment. Due
to the abrupt expansion in area from Plane 2 to Plane
5, it is assumed that there is no conversion of velocity
pressure at Plane 2 to static pressure at Plane 5.
Therefore, it is assumed that Ps2 = Ps5. Measure the
area of Plane 4 for use in calculating Pv4.
4. Measure the dry-bulb and wet-bulb temperatures
at Plane 4 and the dry-bulb temperatures at Planes
3a, 3b, and 5. Determine pb for the general vicinity of
the air conditioning unit. These measurements are
used in determining densities at the planes of interest.
5. Measure the fan speed and motor amps, volts,
and if possible, watts. Record all pertinent motor
nameplate data, including volts (NPV), and full load
amps (FLA). If the motor power output is to be
estimated by using the phase current method
described in Annex K, it is not necessary to measure
motor watts; however, it may be necessary to
disconnect the drive and measure the no load amps
(NLA) if the motor is not operating at or near its full
load point. Refer to Annex K.
6. SEF 1 is due to the effect of insufficient distance
between the fan inlets and the side walls of the fan
cabinet. SEF 2 is attributed to the high degree of
divergence of the transition fitting at the fan outlet.
The effect created by this fitting is considered to be
equivalent to the effect created by having no duct at
the fan outlet. In order to determine the values of the
SEFs, it is necessary to measure the diameter of an
inlet of the fan, the distance between a fan inlet and
a side wall of the fan cabinet, and the outlet area and
blast area of the fan.
7. To calculate the Fan Static Pressure:
Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2
= Ps2 - (Ps1 + Pv1) + SEF 1 + SEF 2
Where:
Ps2 = Ps5
Ps1 + Pv1 = Ps4 + Pv4
Pv4 = (Q4/1096 A4)2 ρ4
Q4 = Q1
= Q3a (ρ3a/ρ1) + Q3b (ρ3b/ρ1)
The calculation of Pv4 is often ignored in instances
similar to this example on the basis that the
calculated value of Pv4 is relatively small and its
omission does not affect the test results significantly.
8. In order to compare the test results to the quoted
fan curve drawn for operation at 1170 rpm and 0.075
lbm/ft3 density, it is necessary to convert the results
to the specified conditions. The basis for the
calculations is described in Section 14.
OBSERVATIONS
SITE MEASUREMENTS
pb = 28.72 in. Hg
td3a = 59°F
td3b = 90°F
td4 = 56°F
td5 = 58°F
Ps4 = -1.75 in. wg
Ps3a = 3.65 in. wg
Ps3b = 3.45 in. wg
Pv3a = 0.60 in. wg
Pv3b = 0.47 in. wg
Ps5 = 3.77 in. wg
N = 1160 rpm
A2 = 18.9 ft2
A3a = 7.2 ft2
A3b = 9.7 ft2
A4 = 93.2 ft2
Blast Area = 13.3 ft2
D1 = 3.92 ft, fan inlet diameter
L
= 2.83 ft
MEASURED MOTOR DATA
Volts = 462, 465, 465
= 464 av
Amps = 82, 81, 83
= 82 av
MOTOR NAMEPLATE DATA
75 hp, 3 phase, 60 hertz
460 volts, 1780 rpm, 90.3 FLA
GENERAL
Fan connected to motor through belt drive.
CALCULATIONS
DENSITIES
For Plane 4 conditions of:
td4 = 56°F
tw4 = 54°F
p4 = pb + (Ps4/13.6)
= 28.72 + (-1.75/13.6)
= 28.59 in. Hg
Field Performance Measurement | 217
Use Figure N.1 in Annex N to obtain ρ4 = 0.0731
lbm/ft3.
It is assumed that ρ1 = ρ4.
⎛ P + 13.6 pb ⎞ ⎛ t d4 + 460 ⎞
ρ5 = ρ 4 ⎜ s5
⎟
⎟⎜
⎝ 13.6 p4 ⎠ ⎝ t d5 + 460 ⎠
⎛ 3.77 + 13.6 × 28.72 ⎞ ⎛ 516 ⎞
= 0.0731⎜
⎟ ⎜ 518 ⎟
13.6 × 28.59
⎝
⎠⎝
⎠
3
= 0.0739 lbm/ft
⎛ P + 13.6 pb ⎞ ⎛ t d4 + 460 ⎞
ρ3a = ρ 4 ⎜ s3a
⎟
⎟⎜
⎝ 13.6 p4
⎠ ⎝ t d3a + 460 ⎠
⎛ 3.65 + 13.6 × 28.72 ⎞ ⎛ 516 ⎞
= 0.0731⎜
⎟ ⎜ 519 ⎟
13.6 × 28.59
⎝
⎠⎝
⎠
= 0.0737 lbm/ft 3
FAN POWER INPUT
Measured amps/FLA = (82/90.3)
= 0.91
= 91%
Annex K indicates that Equation A will provide a
reasonably accurate estimate of motor power output
for a 75 hp motor operating at 91% FLA.
Hmo = 75 (82/90.3) (464/460)
= 68.7 hp
Figure L.1 in Annex L indicates estimated belt drive
loss of 4.3%.
HL = 0.043 Hmo
= 0.043 × 68.7
= 2.95 hp
H
⎛ P + 13.6 pb ⎞ ⎛ t d4 + 460 ⎞
ρ3b = ρ 4 ⎜ s3b
⎟
⎟⎜
⎝ 13.6 p4
⎠ ⎝ t d3b + 460 ⎠
⎛ 3.45 + 13.6 × 28.72 ⎞ ⎛ 516 ⎞
= 0.0731⎜
⎟ ⎜ 550 ⎟
13.6 × 28.59
⎝
⎠⎝
⎠
= 0.0695 lbm/ft 3
= Hmo - HL
= 68.7 - 2.95
= 68.75 hp
SYSTEM EFFECT FACTORS
SEF 1 is due to the effect of insufficient distance
between the fan inlets and the side walls of the fan
plenum. The distance is 2.83 ft, or:
FLOW RATES
V3a = 1096 (Pv3a/ρ3a
= 1096 (0.60/0.0737)0.5
= 3127 fpm
)0.5
V3b = 1096 (Pv3b/ρ3b)0.5
= 1096 (0.47/0.0695)0.5
= 2850 fpm
Q3a = V3aA3a
= 3127 × 7.2
= 22514 cfm
Q3b = V3bA3b
= 2850 × 9.7
= 27645 cfm
Q = Q1
= Q3a (ρ3a/ρ1) + Q3b (ρ3b/ρ1)
= 22514 (0.0737/0.0731) + 27645 (0.0695/0.0731)
= 48982 cfm
Q2 = Q1 (ρ1/ρ2)
= 48982 (0.0731/0.0739)
= 48452 cfm
218 | Field Performance Measurement
(2.83/3.92) = 0.72
= 72%
Of the fan inlet diameter. The area of the fan inlets:
A1 = 2 (π D12/4)
= 2 (π × 3.922/4)
= 24.1 ft2
The fan inlet velocity:
V1 = (Q1/A1)
= (48982/24.1)
= 2032 fpm
AMCA Publication 201-90, Figure 9.11A, indicates
that for a plenum wall spacing of 72% of the fan inlet
diameter System Effect Curve V applies. For 2032
fpm inlet velocity and curve V, Figure 7.1 shows SEF
1 = 0.06 in. wg at 0.075 lbm/ft3 density. At 0.0731
lbm/ft3:
SEF 1 = 0.06 (0.0731/0.075)
= 0.06 in. wg
For SEF 2, AMCA Publication 201-90, Figures 7.1
and 8.3, indicate the following calculations:
V2 = (Q2/A2)
= (48452/18.9)
= 2564 fpm
Blast area ratio = Blast Area/A2
= 13.3/18.9
= 0.70
For a blast area ratio of 0.7 and no duct, Figure 8.3
shows System Effect Curve S applies. For 2564 fpm
velocity and curve S, Figure 7.1 shows SEF 2 = 0.33
in. wg at 0.075 lbm/ft3 density. At 0.0739 lbm/ft3:
SEF 2 = 0.33 (0.0739/0.075)
= 0.33 in. wg
FAN STATIC PRESSURE
Pv4 = (Q4/1096 A4)2 ρ4
Since:
ρ4 = ρ1
Q4 = Q1
Pv4 = (48982/1096 × 93.2)2 0.0731
= 0.02 in. wg
Ps1 + Pv1 = Ps4 + Pv4
= -1.75 + 0.02
= -1.73 in. wg
Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2
= Ps2 - (Ps1 + Pv1) + SEF 1 + SEF 2
= 3.77 - (-1.73) + 0.06 + 0.33
= 5.89 in. wg
CONVERSION TO SPECIFIED CONDITIONS
Qc = 48982 (1170/1160)
= 49404 cfm
Psc = 5.89 (1170/1160)2 (0.075/0.0731)
= 6.15 in. wg
Hc = 65.75 (1170/1160)3 (0.075/0.0731)
= 69.22 hp
Field Performance Measurement | 219
EXAMPLE 4B: CENTRAL STATION AIR CONDITIONING UNIT, FACTORY ASSEMBLED DRAWTHROUGH TYPE
1
PLAN VIEW
RETURN AIR
3
STATIC PRESSURE TAPS
L
5
SEF 1
2
+
OUTSIDE
AIR
+
+
+
FAN SECTION
SIDE VIEW
FILTER SECTION
COIL SECTION
COMMENTS
1. This is a factory assembled, draw-through central
station unit. The subject of the test is the fan section,
which is rated by the manufacturer as an assembly of
the fan and the cabinet in which the fan has been
installed. As a draw-through unit, the performance
ratings for the fan section are based on operation
with the fan outlet ducted. Before proceeding with the
test, it is essential that all dampers--outside air, return
air, mixing box, multizone, face and bypass, or
volume control--be fixed in the positions agreed upon
by all interested parties as being applicable for the
installation. Also, the temperatures of heating and
cooling coils must be kept constant throughout the
test period. It may be necessary to lock out,
disconnect, or otherwise modify automatic control
devices in order to prevent the positions of the
dampers and temperatures of the coils from changing
during the test. Refer to Section 17.4.2 for additional
considerations affecting the test procedure in this
type of installation.
2. Determine Pv3 by using the root mean square of
the velocity pressure measurements made in a
traverse of Plane 3, located near the end of a straight
run of duct, as shown in the diagram. Determine Ps3
by averaging the static pressure measurements
made in the same traverse. This static pressure value
is used to determine the density at the traverse
plane. Procedures for traverses are described in
220 | Field Performance Measurement
Section 9.4. In order to determine the air flow rate, it
is necessary to measure the area of the traverse
plane.
3. Determine Ps1 by averaging the static pressure
measurements made in a traverse of Plane 1. Ps5
may be determined in a similar manner or by
averaging the pressure measurements at each of
four static pressure taps. If it is possible to install
suitable pressure taps, their use is preferred in the
region of the fan outlet. due to the close proximity of
Planes 2 and 5, and the fact that there is no change
in area between the two planes, the conditions which
exist at Plane 5 are assumed to exist at Plane 2.
Measure the area of Plane 1 for use in calculating
Pv1.
4. Measure the dry-bulb and wet-bulb temperatures
at Plane 3 and the dry-bulb temperatures at Planes 1
and 5. Determine pb for the general vicinity of the air
conditioning unit. These measurements are used to
determine densities at the planes of interest.
5. Measure the fan speed and the motor amps, volts,
and if possible, watts. Record all pertinent motor
nameplate data, including volts (NPV), and full load
amps (FLA). If the motor power output is to be
estimated by using the phase current method
described in Annex K, it is not necessary to measure
motor watts; however, it may be necessary to
disconnect the drive and measure the no load amps
(NLA) if the motor is not operating at or near its full
load point. Refer to Annex K.
6. SEF 1 is due to the effect of insufficient length of
duct between the fan outlet and the elbow
downstream of the fan. In order to determine the
value of SEF 1, it is necessary to measure the outlet
area of the fan, A2; the length of the outlet duct, L;
and the blast area of the fan.
MEASURED MOTOR DATA
Volts = 440, 444, 442
= 442 av
Amps = 47.4, 47.7, 48.0
= 47.7 av
MOTOR NAMEPLATE DATA
40 hp, 3 phase, 60 hertz
440 volts, 1770 rpm, 49.7 FLA
7. To calculate the Fan Section Static Pressure:
GENERAL
Ps = Ps2 - Ps1 - Pv1 + SEF 1
Fan connected to motor through belt drive.
CALCULATIONS
Where:
Ps2 = Ps5
DENSITIES
Pv1 = (Q1/1096A1)2 ρ1
For Plane 3 conditions of:
The calculation of Pv1 is often ignored in instances
similar to this example on the basis that the
calculated value of Pv1 is relatively small, and it
omission does not affect the test results significantly.
8. In order to compare the test results to the quoted
fan section curve drawn for operation at 1430 rpm
and 0.075 lbm/ft3 density, it is necessary to convert
the results to the specified conditions. The basis for
the calculations is described in Section 14.
OBSERVATIONS
SITE MEASUREMENTS
pb = 29.27 in. Hg
td1 = 47.5°F
td3 = 49.3°F
tw3 = 47.3°F
td5 = 49°F
Ps1 = -0.847 in. wg
Ps3 = 1.31 in. wg
Pv3 = 0.294 in. wg
Ps5 = 1.39 in. wg
N = 1402 rpm
A1 = 147.2 ft2
A2 = A3 = A5
= 15.42 ft2
Blast Area = 9.4 ft2
L = 2.0 ft, length of outlet duct
td3 = 49.3°F
tw3 = 47.3°F
p3 = pb + (Ps3/13.6)
= 29.27 + (1.31/13.6)
= 29.37 in. Hg
Use Figure N.1 in Annex N to obtain ρ3 = 0.0762
lbm/ft3.
⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞
ρ1 = ρ3 ⎜ s1
⎟⎜
⎟
⎝ 13.6 p3 ⎠ ⎝ t d1 + 460 ⎠
⎛ −0.847 + 13.6 × 29.27 ⎞ ⎛ 509.3 ⎞
= 0.0762 ⎜
⎟ ⎜ 507.5 ⎟
13.6 × 29.37
⎝
⎠⎝
⎠
= 0.0760 lbm/ftt 3
⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞
ρ5 = ρ3 ⎜ s5
⎟⎜
⎟
⎝ 13.6 p3 ⎠ ⎝ t d5 + 460 ⎠
⎛ 1.39 + 13.6 × 29.27 ⎞ ⎛ 509.3 ⎞
= 0.0762 ⎜
⎟ ⎜ 509 ⎟
13.6 × 29.37
⎝
⎠⎝
⎠
= 0.0763 lbm/ft 3
It is assumed ρ2 = ρ5.
FLOW RATES
V3 = 1096 (Pv3/ρ3)0.5
= 1096 (0.294/0.0762)0.5
= 2153 fpm
Field Performance Measurement | 221
Q3 = V3A3
= 2153 × 15.42
= 33199 cfm
Q = Q1
= Q3 (ρ3/ρ1)
= 33199 (0.0762/0.0760)
= 33286 cfm
Q2 = Q5
= Q3 (ρ3/ρ5)
= 33199 (0.0762/0.0763)
= 33155 cfm
FAN POWER INPUT
Measured amps/FLA = (47.7/49.7)
= 0.96
= 96%
Annex K indicates that Equation A will provide a
reasonably accurate estimate of motor power output
for a 40 hp motor operating at 96% FLA.
Hmo = 40 (47.7/49.7) (442/440)
= 38.6 hp
Figure L.1 in Annex L indicates estimated belt drive
loss of 4.5%.
HL = 0.045 Hmo
= 0.045 × 38.6
= 1.74 hp
H
= Hmo - HL
= 38.6 - 1.74
= 36.86 hp
For velocities of 2500 fpm or less, the 100% effective
outlet duct length is 2.5 duct diameters:
= 2.5 × 4.43
= 11.1 ft
The length of the outlet duct in % effective duct
length:
= (L/11.1) 100
= (2.0/11.1) 100
= 18%
Blast area ratio = Blast Area/A2
= 9.4/15.42
= 0.61
For a blast area ratio of 0.6, 18% effective duct length
and elbow position A, Figure 8.5 shows System
Effect Curve R applies. For 2150 fpm velocity and
curve R, Figure 7.1 shows SEF 1 = 0.34 in. wg at
0.075 lbm/ft3 density. At 0.0762 lbm/ft3:
SEF 1 = 0.34 (0.0762/0.075)
= 0.35 in. wg
FAN SECTION STATIC PRESSURE
Pv1 = (Q1/1096 A1)2 ρ1
= (33286/1096 × 147.2)2 0.0760
= 0.003 in. wg
It is assumed that Ps2 = Ps5
Ps = Ps2 - Ps1 - Pv1 + SEF 1
= 1.39 - (-0.847) - 0.003 + 0.35
= 2.58 in. wg
SYSTEM EFFECT FACTOR
CONVERSION TO SPECIFIED CONDITIONS
To determine SEF 1, AMCA Publication 201-90,
Figures 7.1 and 8.5, indicate the following
calculations:
Qc = 33286 (1430/1402)
= 33951 cfm
V2 = (Q2/A2)
= (33155/15.42)
= 2150 fpm
Duct diameter equivalent to the fan outlet area:
De2 = (4 A2/π)0.5
= (4 × 15.42/π)0.5
= 4.43 ft
222 | Field Performance Measurement
Psc = 2.58 (1430/1402)2 (0.075/0.0760)
= 2.65 in. wg
Hc = 36.86 (1430/1402)3 (0.075/0.0760)
= 38.60 hp
EXAMPLE 4C: PACKAGED AIR-CONDITIONING UNIT
3
2
L
SEF 1
PLAN VIEW
4 1
INLET PLENUM
FILTERS
FANS
5
+
+
COOLING COIL
SIDE VIEW
COMMENTS
1. The subject of the test in this example is the air
conditioning unit assembly. This assembly does not
include the inlet plenum. The performance ratings for
the unit assembly are based on operation with the
outlets of the fans ducted. Before proceeding with the
test, it is essential that all system dampers be fixed in
the positions agreed upon by all interested parties as
being applicable for the installation. Also, the
temperature of the cooling coil must be kept constant
throughout the test period. It may be necessary to
lock out, disconnect or otherwise modify automatic
control devices in order to prevent the positions of the
dampers and the temperature of the coil from
changing during the test. Refer to Section 17.4.1 for
additional considerations affecting the test procedure
in this type of installation.
3. Ps4 may be determined by averaging the pressure
measurements at each of four static pressure taps or
by averaging the static pressure measurements
made in a Pitot-static tube traverse of Plane 4. Ps5 is
determined in a similar manner. However, if it is
possible to install suitable static pressure taps, their
use is preferred in the regions of the outlets of the
fans. Due to the close proximity of Planes 1 and 4
and the fact that there is no change in area between
the two planes, the conditions which exist at Plane 4
are assumed to exist at Plane 1. Although Plane 5 is
greater in area that Plane 2, the degree of divergence
is relatively small. Therefore, Ps2 will be calculated
based on Ps5 and the assumption that there is no
change in total pressure from Plane 2 to Plane 5.
2. Determine Pv3 by using the root mean square of
the velocity pressure measurements made in a
traverse of Plane 3, located near the end of a straight
run of duct, as shown in the diagram. Determine Ps3
by averaging the static pressure measurements
made in the same traverse. This static pressure value
is used to determine the density at the traverse
plane. Procedures for traverses are described in
Section 9.4. in order to determine the air flow rate, it
is necessary to measure the area of the traverse
plane.
4. Measure the dry-bulb and wet-bulb temperatures
at Plane 4 and the dry-bulb temperatures at Planes 3
and 5. In this example, the cooling medium, normally
circulated in the coil was shut off in order to maintain
constant air temperatures during the test. In order to
account for water vapor which may have been added
to the air as a result of evaporation of moisture
previously condensed on the coil, the wet-bulb
temperature at Plane 3 was measured. Determine pb
for the general vicinity of the air conditioning unit.
These measurements are used in determining
densities at the planes of interest.
Field Performance Measurement | 223
5. Measure the fan speed and the motor amps, volts,
and if possible, watts. Record all pertinent motor
nameplate data including volts (NPV), and full load
amps (FLA). If the motor power output is to be
estimated by using the phase current method
described in Annex K, it is not necessary to measure
motor watts; however, it may be necessary to
disconnect the drive and measure the no load amps
(NLA) if the motor is not operating at or near its full
load point. Refer to Annex K.
A1 = A4
= 31.7 ft2
A2 = 11.5 ft2
A3 = 16.4 ft2
A5 = 14.3 ft2
Blast Area = 4.0 ft2 per fan
L = 2.0 ft, length of outlet duct
6. Although an elbow is located shortly downstream
of the fans, SEF 1 is judged to be more closely
characterized as the effect due to insufficient lengths
of duct on the outlets of the fans. In order to
determine the value of SEF 1, it is necessary to
measure the outlet area and the blast area of one of
the fans and the length, L, of its outlet duct.
Volts = 460, 455, 465
= 460 av
Amps = 38.2, 38, 37.9
= 38.0 av
7. To calculate the static pressure for the unit
assembly:
MEASURED MOTOR DATA
MOTOR NAMEPLATE DATA
25 hp, 3 phase, 60 hertz
460 volts, 1760 rpm, 39.5 FLA
GENERAL
Ps = Ps2 - Ps1 - Pv1 + SEF 1
Fans connected to motor through belt drive.
Where:
CALCULATIONS
Ps1 = Ps4
DENSITIES
Pv1 = (Q1/1096A1)2 ρ1
Ps2 = Ps5 + Pv5 - Pv2
Pv2 and Pv5 are calculated in manners similar to the
calculation of Pv1.
8. In order to compare the test results to the quoted
unit assembly curve drawn for operation at 1050 rpm
and 0.075 lbm/ft3 density, it is necessary to convert
the results to the specified conditions. The basis for
the calculations is described in Section 14.
OBSERVATIONS
SITE MEASUREMENTS
pb = 29.65 in. Hg
td3 = 75.0°F
tw3 = 59.5°F
td4 = 72.5°F
tw4 = 58.5°F
td5 = 74.5°F
Ps3 = 2.02 in. wg
Pv3 = 0.35 in. wg
Ps4 = -0.32 in. wg
Ps5 = 2.11 in. wg
N = 1025 rpm
224 | Field Performance Measurement
For Plane 3 conditions of:
td3 = 75.0°F
tw3 = 59.5°F
p3 = pb + (Ps3/13.6)
= 29.65 + (2.03/13.6)
= 29.80 in. Hg
Use Figure N.1 in Annex N to obtain ρ3 = 0.0736
lbm/ft3.
For Plane 4 conditions of:
td4 = 72.5°F
tw4 = 58.5°F
p4 = pb + (Ps4/13.5)
= 29.65 + (-0.32/13.6)
= 29.63 in. Hg
Use Figure N.1 in Annex N to obtain ρ4 = 0.0735
lbm/ft3.
It is assumed that ρ1 = ρ4.
⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞
ρ5 = ρ3 ⎜ s5
⎟⎜
⎟
⎝ 13.6 p3 ⎠ ⎝ t d5 + 460 ⎠
⎛ 2.11 + 13.6 × 29.65 ⎞ ⎛ 535 ⎞
= 0.0736 ⎜
⎟ ⎜ 534.5 ⎟
13.6 × 29.80
⎝
⎠⎝
⎠
= 0.0737 lbm/ft 3
SYSTEM EFFECT FACTOR
To determine SEF 1, AMCA Publication 201-90,
Figures 7.1 and 8.3, indicate the following
calculations:
It is assumed ρ2 = ρ5.
V2 = (Q2/A2)
= (39143/11.5)
= 3404 fpm
FLOW RATES
Duct diameter equivalent to the outlet area of one fan:
V3 = 1096 (Pv3/ρ3)0.5
= 1096 (0.35/0.0736)0.5
= 2390 fpm
De2 = (4A2/2π)0.5
= (4 × 11.5/2π)0.5
= 2.71 ft
Q3 = V3A3
= 2390 × 16.4
= 39196 cfm
Figure 8.3 shows that for velocities over 2500 fpm,
100% effective duct length is one diameter for every
1000 fpm:
Q2 = Q5
= Q3 (ρ3/ρ5)
= 39196 (0.0736/0.0737)
= 39143 cfm
Q = Q1 = Q4
= Q3 (ρ3/ρ4)
= 39196 (0.0736/0.0735)
= 39249 cfm
FAN POWER INPUT
Measured amps/FLA = (38.0/39.5)
= 0.96
= 96%
= De2 (V2/1000)
= 2.71 (3404/1000)
= 9.22 ft
L in % effective duct length:
= (L/9.22) 100
= (2.0/9.22) 100
= 22%
Blast area ratio = Blast area/A2
= (2 × 4.0)/11.5
= 0.70
Annex K indicates that Equation A will provide a
reasonably accurate estimate of motor power output
for a 25 hp motor operating at 96% FLA.
For a blast area ratio of 0.7, and 22% effective duct
length Figure 8.3 shows System Effect Curve W
applies. For 3404 fpm velocity and curve W, Figure
7.1 shows SEF 1 = 0.13 in. wg at 0.075 lbm/ft3
density. At 0.0737 lbm/ft3:
Hmo = 25 (38.0/39.5) (460/460)
= 24.1 hp
SEF 1 = 0.13 (0.0737/0.075)
= 0.13 in. wg
Figure L.1 in Annex L indicates estimated belt drive
loss of 4.8%.
STATIC PRESSURE OF UNIT
HL = 0.048 Hmo
= 0.048 × 24.1
= 1.2 hp
H = Hmo - HL
= 24.1 - 1.2
= 22.9 hp
Pv5 = (Q5/1096 A5)2 ρ5
= (39143/1096 × 14.3)2 0.0737
= 0.46 in. wg
Pv2 = (Q2/1096 A2)2 ρ2
= (39143/1096 × 11.5)2 0.0737
= 0.71 in. wg
Field Performance Measurement | 225
Ps2 + Pv2 = Ps5 + Pv5
Ps2 = Ps5 + Pv5 - Pv2
= 2.11 + 0.46 - 0.71
= 1.86 in. wg
Pv1 = (Q1/1096 A1)2 ρ1
= (39249/1096 × 31.7)2 0.0735
= 0.09 in. wg
Ps = Ps2 - Ps1 - Pv1 + SEF 1
= 1.86 - (-0.32) - 0.09 + 0.13
= 2.22 in. wg
CONVERSION TO SPECIFIED CONDITIONS
Qc = 39249 (1050/1025)
= 40206 cfm
Psc = 2.22 (1050/1025)2 (0.075/0.0735)
= 2.38 in. wg
Hc = 22.9 (1050/1025)3 (0.075/0.0735)
= 25.1 hp
226 | Field Performance Measurement
EXAMPLE 4D: PACKAGED AIR-CONDITIONING UNIT
3a
3b
PLAN VIEW
STATIC PRESSURE TAPS
2
5
L
FILTER SECTION
1
SEF 1
+
+
INLET LOUVER
HEATING COIL
SIDE VIEW
COMMENTS
1. The subject of the test in this example is the air
conditioning unit assembly. This assembly includes
the filter section and the inlet louver. The
performance ratings for the unit assembly are based
on operation with the outlets of the fans ducted.
Before proceeding with the test, it is essential that all
system dampers be fixed in the positions agreed
upon by all interested parties as being applicable for
the installation. Also, the temperature of the heating
coil must be kept constant throughout the test period.
It may be necessary to lock out, disconnect or
otherwise modify automatic control devices in order
to prevent the positions of the dampers and the
temperature of the coil from changing during the test.
Refer to Section 17.5.1 for additional considerations
affecting the test procedure in this type of installation.
2. Normally, velocity pressure measurements would
be made in a single plane, located in a duct common
to all branches. In this example, a measurement
plane which provides a satisfactory velocity profile
cannot be located upstream of the fans or between
the point of connection of the branch ducts and the
outlets of the fans. The alternative, as indicated in the
diagram, is to make a velocity pressure
measurement traverse in each of two branches. the
velocity pressure for reach branch is determined by
using the root mean square of the velocity pressure
measurements made in the traverse. The static
pressure at each traverse plane is determined by
using the root mean square of the velocity
measurement traverse in each of two branches. The
velocity pressure for each branch is determined by
using the root mean square of the velocity pressure
measurements made in the traverse. The static
pressure at each traverse plane is determined by
averaging the static pressure measurements made in
the same traverse. These static pressure values are
used in determining the densities at the traverse
planes. Procedures for traverses are described in
Section 9.4. In order to determine the air flow rates, it
is necessary to measure the area of each traverse
plane.
3. Determine Ps5 by averaging the pressure
measurements at each of four static pressure taps
located in the duct fitting at the outlets of the fans.
The conditions which exist at Plane 5, including the
static pressure, are assumed to exist at Plane 2,
based on their close proximity and the fact that there
is no change in area between the two planes. In
situations such as this example, it is important to be
certain that all pressure measurements are referred
to the same atmospheric pressure.
4. Measure the dry-bulb and wet-bulb temperatures
at Plane 1 and the dry-bulb temperatures at Planes
3a, 3b, and 5. Determine pb for the general vicinity of
Field Performance Measurement | 227
the air conditioning unit. These measurements are
used to determine densities at the planes of interest.
5. Measure the fan speed and the motor amps, volts,
and if possible, watts. Record all pertinent motor
nameplate data, including volts (NPV), and full load
amps (FLA). If the motor power output is to be
estimated by using the phase current method
described in Annex K, it is not necessary to measure
motor watts; however, it may be necessary to
disconnect the drive and measure the no load amps
(NLA) if the motor is not operating at or near its full
load point. Motor performance data, supplied by the
motor manufacturer, are used in the determination of
motor power output in this example.
6. SEF 1 is due to the effect of insufficient length of
duct between the outlets of the fans and the elbow
downstream of the fans. In order to determine the
value of SEF 1, it is necessary to measure the outlet
area and the blast area of one of the fans and the
length of the duct, L, between the fan and the elbow.
7. The sum of the static pressure, Ps1, and velocity
pressure, Pv1, at the inlet to the unit assembly is
considered to be equal to the sum of the static
pressure, Psx, and velocity pressure, Pvx, at a point
sufficiently distant from the inlet as to be in still air. At
this point, the static pressure is zero, and the velocity
pressure in still air is zero.
Ps1 + Pv1 = Psx + Pvx = 0
This consideration, which is the same as that used in
the methods for testing this type of unit for
performance rating purposes, charges to the unit
losses incurred in accelerating the air into its inlet and
eliminates the inaccuracies which arise in any
attempt to measure the velocity pressure and static
pressure at the inlet. To calculate the static pressure
for the unit assembly:
Ps = Ps2 - Ps1 - Pv1 + SEF 1
= Ps2 - (Ps1 + Pv1) + SEF 1
Since:
Ps1 + Pv1 = 0
Ps = Ps2 + SEF 1
Where:
Ps2 = Ps5
8. In order to compare the test results to the quoted
performance curve for the packaged unit drawn for
operation at 1720 rpm and 0.075 lbm/ft3 density, it is
necessary to convert the results to the specified
conditions. The basis for the calculations is described
in Section 14.
OBSERVATIONS
SITE MEASUREMENTS
pb = 29.65 in. Hg
td1 = 72°F
tw1 = 61°F
td5 = 85°F
td3a = 82.5°F
td3b = 83°F
Ps5 = 1.25 in. wg
Ps3a = 1.15 in. wg
Ps3b = 1.22 in. wg
Pv3a = 0.56 in. wg
Pv3b = 0.60 in. wg
N = 1710 rpm
A2 = A5
= 5.64 ft2
A3a = 3.1 ft2
A3b = 2.2 ft2
Blast Area = 2.5 ft2 per fan
L = 0.96 ft, length of outlet duct
MEASURED MOTOR DATA
Volts = 460, 458, 462
= 460 av
Amps = 10.0, 10.0, 9.8
= 9.9 av
MOTOR NAMEPLATE DATA
10 hp, 3 phase, 60 hertz
460 volts, 1750 rpm, 13.5 FLA
GENERAL
Fans connected to motor through belt drive. The
following motor performance data was supplied by
the motor manufacturer:
Motor Efficiency:
82.5% at 1/2 load
84.5% at 3/4 load
84.5% at full load
Power Factor = 0.85
228 | Field Performance Measurement
DENSITIES
For Plane 1 conditions of:
td1 = 72°F
tw1 = 61°F
p1 = pb
= 29.65 in. Hg
Use Figure N.1 in Annex N to obtain ρ1 = 0.0735
lbm/ft3.
⎛ P + 13.6 pb ⎞ ⎛ t d1 + 460 ⎞
ρ5 = ρ1 ⎜ s5
⎟
⎟⎜
⎝ 13.6 p1 ⎠ ⎝ t d5 + 460 ⎠
⎛ 1.25 + 13.6 × 29.65 ⎞ ⎛ 532 ⎞
= 0.0735 ⎜
⎟ ⎜ 545 ⎟
13.6 × 29.65
⎝
⎠⎝
⎠
= 0.0720 lbm/ft 3
It is assumed that ρ2 = ρ5
⎛ P + 13.6 pb ⎞ ⎛ t d1 + 460 ⎞
ρ3a = ρ1 ⎜ s3a
⎟
⎟⎜
⎝ 13.6 p1
⎠ ⎝ t d3a + 460 ⎠
⎛ 1.15 + 13.6 × 29.65 ⎞ ⎛ 532 ⎞
= 0.0735 ⎜
⎟ ⎜ 542.5 ⎟
13.6 × 29.65
⎝
⎠⎝
⎠
3
= 0.0723 lbm/ft
⎛ P + 13.6 pb ⎞ ⎛ t d1 + 460 ⎞
ρ3b = ρ1 ⎜ s3b
⎟
⎟⎜
⎝ 13.6 p1
⎠ ⎝ t d3b + 460 ⎠
⎛ 1.22 + 13.6 × 29.65 ⎞ ⎛ 532 ⎞
= 0.0735 ⎜
⎟ ⎜ 543 ⎟
13.6 × 29.65
⎝
⎠⎝
⎠
= 0.0722 lbm/ft 3
FLOW RATES
V3a = 1096 (Pv3a/ρ3a)0.5
= 1096 (0.56/0.0723)0.5
= 3050 fpm
V3b = 1096 (Pv3b/ρ3b
= 1096 (0.60/0.0722)0.5
= 3159 fpm
)0.5
Q3a = V3aA3a
= 3050 × 3.1
= 9455 cfm
Q3b = V3bA3b
= 3159 × 2.2
= 6950 cfm
Q = Q1
= Q3a (ρ3a/ρ1) + Q3b (ρ3b/ρ1)
= 9455 (0.0723/0.0735) + 6950 (0.0722/0.0735)
= 16128 cfm
Q2 = Q5
= Q1 (ρ1/ρ5)
= 16128 (0.0735/0.0720)
= 16464 cfm
FAN POWER INPUT
Measured amps/FLA = (9.9/13.5)
= 0.73
= 73%
The data supplied by the motor manufacturer indicate
power factor of 0.85 and motor efficiency of 84.5% for
the motor operating at 73% FLA. Using the
appropriate equation in Section 10.2.2:
Hmo = (3)0.5 × 9.9 × 460 × 0.85 × 0.845/746
= 7.59 hp
Figure L.1 in Annex L indicates estimated belt drive
loss of 5.6%.
HL = 0.056 Hmo
= 0.056 × 7.59
= 0.43 hp
H
= Hmo - HL
= 7.59 - 0.43
= 7.16 hp
SYSTEM EFFECT FACTOR
SEF 1 is due to the effect of insufficient lengths of
duct between the outlets of the fans and the elbow
downstream of the fans. AMCA Publication 201-90,
Figures 7.1, 8.1, and 8.5 indicate the following
calculations:
V2 = (Q2/A2)
= (16464/5.64)
= 2919 fpm
Duct diameter equivalent to the outlet area of one
fan:
De2 = (4A2/2π)0.5
= (4 × 5.64/2π)0.5
= 1.89 ft
Figure 8.1 shows that for velocities over 2500 fpm
100% effective duct length is one diameter for every
1000 fpm:
Field Performance Measurement | 229
= De2 (V2/1000)
= 1.89 (2919/1000)
= 17%
L, in % effective duct length:
= (L/5.52) 100
= (0.96/5.52) 100
= 17%
Blast area ratio = Blast Area/A2
= (2 × 2.5)/5.64
= 0.89
For a blast area ratio of 0.89, 17% effective duct
length and elbow position C, Figure 8.5 shows
System Effect Curve S applies. For 2919 fpm velocity
and curve S, Figure 7.1 shows SEF 1 = 0.43 in. wg at
0.075 lbm/ft3 density. At 0.0720 lbm/ft3:
SEF 1 = 0.43 (0.0720/0.075)
= 0.41 in. wg
STATIC PRESSURE OF UNIT
Ps2 = Ps5
= 1.25 in. wg
Ps = Ps2 + SEF 1
= 1.25 + 0.41
= 1.66 in. wg
CONVERSION TO SPECIFIED CONDITIONS
Qc = 16128 (1720/1710)
= 16222 cfm
Psc = 1.66 (1720/1710)2 (0.075/0.0735)
= 1.71 in. wg
Hc = 7.16 (1720/1710)3 (0.075/0.0735)
= 7.44 hp
230 | Field Performance Measurement
EXAMPLE 4E: CENTRAL STATION AIR CONDITIONING UNIT, FACTORY ASSEMBLED BLOWTHROUGH TYPE
2
1
STATIC PRESSURE TAPS
5
PLAN VIEW
RETURN
AIR
3b
SPRAY
SECTION
3a
HEATING COIL
+
+
+
OUTSIDE
AIR
+
+
+
+
FILTER SECTION
+
FAN SECTION
COOLING COIL
SIDE VIEW
COMMENTS
1. This is a factory assembled, blow-through central
station unit. The subject of the test is the fan section,
which is rated by the manufacturer as an assembly of
the fan and the cabinet in which the fan has been
installed. As a blow-through unit, the performance
ratings for the fan section are based on operation
without the fan outlet ducted. Before proceeding with
the test, it is essential that all dampers (outside air,
return air, mixing box, multizone, face and bypass, or
volume control) be fixed in the positions agreed upon
by all interested parties as being applicable for the
installation. Also, the temperatures of heating and
cooling coils must be kept constant throughout the
test period. It may be necessary to lock out,
disconnect, or otherwise modify automatic control
devices in order to prevent the positions of the
dampers and temperatures of the coils from changing
during the test. In instances in which a cooling coil is
located between a velocity pressure traverse plane
and the fan, as in this example, the flow of the cooling
medium should be stopped or its temperature raised
to a level sufficient to prevent condensation on the
cooling coil, otherwise the moisture condensed will
not be properly taken into account in the
determination of fan air flow rate. Refer to Section
17.5.2 for additional considerations affecting the test
procedure in this type of installation.
2. Normally, velocity pressure measurements would
be made in a single plane, located in a duct common
to all branches. In this example, a measurement
plane which provides a satisfactory velocity profile
cannot be located upstream of the fan or between the
point of connection of the branch ducts and the fan
outlet. The alternative, as indicated in the diagram, is
to make a velocity pressure measurement traverse in
each branch. The velocity pressure for each branch
is determined by using the root mean square of the
velocity pressure measurements made in the
traverse. The static pressure at each traverse plane
is determined by averaging the static pressure
measurements made in the same traverse. These
static pressure values are used in determining the
densities at the traverse plane. Procedures for
traverses are described in Section 9.4. In order to
determine the air flow rates it is necessary to
measure the area of each traverse plane.
3. Determine Ps1 by averaging the static pressure
measurements made in a traverse of Plane 1. Ps5
may be determined in a similar manner or by
averaging the pressure measurements at each of
four static pressure taps. If it is possible to install
suitable pressure taps, their use is preferred in the
regions of the fan outlet. Due to the abrupt expansion
in area from Plane 2 to Plane 5, it is assumed that
there is no conversion of velocity pressure at Plane 2
to static pressure at Plane 5. Therefore, it is assumed
Field Performance Measurement | 231
that Ps2 = Ps5. Measure the area of Plane 1 for use in
calculating Pv1.
4. Measure the dry-bulb and wet-bulb temperatures
at Planes 1, 3a, and 3b. Determine pb for the general
vicinity of the air conditioning unit. These
measurements are used to determine densities at the
planes of interest. The measurements of additional
wet-bulb temperatures were made in this example in
order to provide data which may be used to
determine whether the moisture content of the air
changed between Plane 1 and Planes 3a and 3b.
5. Measure the fan speed and the motor amps, volts,
and if possible, watts. Record all pertinent motor
nameplate data, including volts (NPV), and full load
amps (FLA). If the motor power output is to be
estimated by using the phase current method
described in Annex K, it is not necessary to measure
motor watts; however, it may be necessary to
disconnect the drive and measure the no load amps
(NLA) if the motor is not operating at or near its full
load point. Refer to Annex K.
tw3a = 71.5°F
td3b = 60°F
tw3b = 58°F
Ps1 = -2.43 in. wg
Ps5 = 6.55 in. wg
Ps3a = 5.35 in. wg
Ps3b = 5.1 in. wg
Pv3a = 0.53 in. wg
Pv3b = 0.60 in. wg
N = 1695 rpm
A1 = 68.9 ft2
A3a = 5.37 ft2
A3b = 6.78 ft2
MEASURED MOTOR DATA
Volts = 570, 575, 565
= 570 av
Amps = 81.5, 82.5, 81
= 81.7
NLA = 19
MOTOR NAMEPLATE DATA
6. Since the performance ratings for the fan section
are based on operation without the fan outlet ducted,
an SEF does not apply for the unducted position.
100 hp, 3 phase, 60 hertz
575 volts, 1790 rpm, 95 FLA
7. To calculate the Fan Section Static Pressure:
GENERAL
Ps = Ps2 - Ps1 - Pv1
Fan connected to motor through belt drive.
Where:
CALCULATIONS
Ps2 = Ps5
Pv1 = (Q1/1096 A1)2 ρ1
Q1 = Q3a (ρ3a/ρ1) + Q3b (ρ3b/ρ1)
DENSITIES
The calculation of Pv1 is often ignored in instances
similar to this example on the basis that the
calculated value of Pv1 is relatively small, and its
omission does not affect the test results significantly.
td1 = 65°F
tw1 = 60°F
8. In order to compare the test results to the quoted
fan section curve drawn for operation at 1650 rpm
and 0.075 lbm/ft3 density, it is necessary to convert
the results to the specified conditions. The basis for
the calculations is described in Section 14.
For Plane 1 conditions of:
p1 = pb + (Ps1/13.6)
= 28.85 + (-2.43/13.6)
= 28.67 in. Hg
Use Figure N.1 in Annex N to obtain ρ1 = 0.0720
lbm/ft3.
For Plane 3a conditions of:
OBSERVATIONS
SITE MEASUREMENTS
pb = 28.85 in. Hg
td1 = 65°F
tw1 = 60°F
td3a = 100°F
232 | Field Performance Measurement
td3a = 100°F
tw3a = 71.5°F
p3a = pb + (Ps3a/13.6)
= 28.85 + (5.35/13.6)
= 29.24 in. Hg
Use Figure N.1 in Annex N to obtain ρ1 = 0.0720
lbm/ft3.
Hmo
For Plane 3b conditions of:
Reference to Figure L.1 in Annex L indicates
estimated belt drive loss of 4.2%.
td3b = 60°F
tw3b = 58°F
p3b = pb + (Ps3b/13.6)
= 28.85 + (5.1/13.6)
= 29.23 in. Hg
Use Figure N.1 in Annex N to obtain ρ3b = 0.0741
lbm/ft3.
FLOW RATES
V3a = 1096 (Pv3a/ρ3a)0.5
= 1096 (0.53/0.0691)0.5
= 3035 fpm
V3b = 1096 (Pv3b/ρ3b)0.5
= 1096 (0.60/0.0741)0.5
= 3119 fpm
Q3a = V3aA3a
= 3035 × 5.37
= 16298 fpm
Q3b = V3bA3b
= 3119 × 6.78
= 21147 cfm
Q = Q1
= Q3a (ρ3a/ρ1) + Q3b (ρ3b/ρ1)
= 16298 (0.0691/0.0720) + 21147 (0.0741/0.0720)
= 37405 cfm
= (85.3 + 81.8)/2
= 83.6 hp
HL = 0.042 Hmo
= 0.042 × 83.6
= 3.5 hp
H = Hmo - HL
= 83.6 - 3.5
= 80.1 hp
FAN SECTION STATIC PRESSURE
Pv1 = (Q1/1096 A1)2 ρ1
= (37405/1096 × 68.9)2 0.0720
= 0.02 in. wg
It is assumed that Ps2 = Ps5
Ps = Ps2 - Ps1 - Pv1
= 6.55 - (-2.43) - 0.02
= 8.96 in. wg
CONVERSION TO SPECIFIED CONDITIONS
Qc = 37405 (1650/1695)
= 36412 cfm
Psc = 8.96 (1650/1695)2 (0.075/0.0720)
= 8.84 in. wg
Hc = 80.1 (1650/1695)3 90.075/0.0720)
= 77.0 hp
FAN POWER INPUT
Measured amps/FLA = (81.7/95)
= 0.86
= 86%
Annex K indicates that the average of the results of
Equation A and Equation B will provide a reasonably
accurate estimate of motor power output for a 100 hp
motor operating at 86% of FLA.
Eqn. A = 100 (81.7/95) (570/575)
= 85.3 hp
Eqn. B = 100 [(81.7 - 19)/(95 - 19)] (570/575)
= 81.8 hp
Field Performance Measurement | 233
EXAMPLE 5A: FREE INLET, FREE OUTLET ROOF VENTILATOR
2
1
3
2 De
TEMPORARY DUCT
WITH SQUARE
CROSS-SECTION,
De = EQUIVALENT
DIAMETER OF DUCT
1.5 De
COMMENTS
1. The subject of the test in this example is the roof
ventilator assembly. Before proceeding with the test,
refer to Section 17.4 for considerations affecting the
test procedure in this type of installation.
2. Determine Pv3 by using the root mean square of
the velocity pressure measurements made in a
traverse of Plane 3, located in the duct which has
been installed on the inlet side of the ventilator.
Determine Ps3 by averaging the static pressure
measurements made in the same traverse.
Procedures for traverses are described in Section
9.4. Measure the area of the traverse plane, A3,
which is located at the tip of the Pitot-static tube. The
duct, temporarily installed for purposes of the test, is
square in cross-section. Its cross-sectional dimensions
were selected as the maximum permissible for its
installation into the opening in the ventilator mounting
curb. The length of the duct is twice its equivalent
diameter and the entrance to the duct is flared in oder
to reduce inlet losses. The installation of a duct of this
size and cross-sectional configuration is judged as
creating no significant effect on the performance of
the ventilator in this example.
3. Ps2, the static pressure at the outlet of the
ventilator, is zero gauge pressure, referred to the
atmospheric pressure in the region of the ventilator
outlet. In situations such as this example, the air may
234 | Field Performance Measurement
be discharging from the ventilator into a region in
which the atmospheric pressure is somewhat
different from that to which all other pressure
measurements are referred. When this possibility
exists, it is essential that the static pressure in the
region of the discharging air be measured, referred to
the same atmospheric pressure as used in all other
pressure measurements. In this example, Ps2 was
measured, referred to the same atmospheric pressure
as in the static pressure measurements made at
Plane 3.
4. Measure the dry-bulb and wet-bulb temperatures
at the velocity traverse plane. Determine pb for the
general vicinity of the ventilator. These measurements
are used to determine densities at the planes of interest.
5. Measure the fan speed and the motor amps and
volts. Record all pertinent motor nameplate data. For
the horsepower rating of the motor in this example, it
is recommended that the fan power input be
determined by using the measured watts input to the
motor and motor performance data, obtained from
the motor manufacturer.
6. To calculate the Fan Static Pressure:
Ps = Ps2 - Ps1 - Pv1
= Ps2 - (Ps1 + Pv1)
Where:
FLOW RATE
Ps1 + Pv1 = Ps3 + Pv3
V3 = 1096 (Pv3/ρ3)0.5
= 1096 (0.077/0.0727)0.5
= 1128 fpm
7. In order to compare the test results to the quoted
fan curve drawn for operation at 1180 rpm and 0.075
lbm/ft3 density, it is necessary to convert the results
to the specified conditions. The basis for the
calculations is described in Section 14.
OBSERVATIONS
SITE MEASUREMENTS
pb = 29.37 in. Hg
td3 = 73.5°F
tw3 = 58.1°F
Ps2 = 0.037 in. wg
Ps3 = -0.085 in. wg
Pv3 = 0.077 in. wg
N = 1177 rpm
A3 = 5.58 ft2
Q = Q1 = Q3
= V3A3
= 1128 × 5.58
= 6294 cfm
FAN POWER INPUT
At the measured power input value of 755 watts, the
data supplied by the motor manufacturer indicate
efficiency of 61% for the motor.
Hmo = (755 × 0.61)/746
= 0.62 hp
Since the fan is direct connected to the motor, there
is no drive loss, and:
MEASURED MOTOR DATA
H = Hmo
= 0.62 hp
Volts = 235, 230, 230
= 232 av
Watts = 755
FAN STATIC PRESSURE
MOTOR NAMEPLATE DATA
1 hp, 3 phase, 60 hertz
230 volts, 1175 rpm, 3.6 FLA
Ps1 + Pv1 = Ps3 + Pv3
= -0.085 + 0.077
= -0.008 in. wg
Ps = Ps2 - (Ps1 + Pv1)
= 0.037 - (-0.008)
= 0.045 in. wg
General
CONVERSION TO SPECIFIED CONDITIONS
Fan direct connected to motor. Motor efficiency data
supplied by motor manufacturer.
Qc = 6294 (1180/1177)
= 6310 cfm
CALCULATIONS
DENSITIES
For Plane 3 conditions of:
td3 = 73.5°F
tw3 = 58.1°F
Psc = 0.045 (1180/1177)2 (0.075/0.0727)
= 0.047 in. wg
Hc = 0.62 (1180/1177)3 (0.075/0.0727)
= 0.64 hp
p3 = pb + (Ps3/13.6)
= 29.37 + (-0.085/13.6)
= 29.36 in. Hg
Use Figure N.1 in Annex N to obtain ρ3 = 0.0727
lbm/ft3.
It is assumed that ρ1 = ρ3.
Field Performance Measurement | 235
EXAMPLE 5B: FREE INLET, FREE OUTLET PROPELLER FAN
2 De
2
3
1.5 De
D2
TEMPORARY DUCT
WITH SQUARE
CROSS-SECTION,
De = EQUIVALENT
DIAMETER OF DUCT
COMMENTS
1. The subject of the test in this example is the
propeller fan assembly. Before proceeding with the
test, refer to Section 17.4 for considerations affecting
the test procedure in this type of installation.
2. Determine Pv3 by using the root mean square of
the velocity pressure measurements made in a
traverse of Plane 3, located in the duct which has
been installed on the inlet side of the fan. Determine
Ps3 by averaging the static pressure measurements
made in the same traverse. Procedures for traverses
are described in Section 9.4. Measure the area of the
traverse plane, A3, which is located at the tip of the
Pitot-static tube. The duct, temporarily installed for
purposes of the test, is square in cross-section, with
side dimension of 1.5 D2. The shape and area of the
duct cross-section were selected on the basis of
minimizing the effect of the duct on the performance
of the fan while providing velocity pressure readings
of measurable magnitudes. The length of the duct is
twice its equivalent diameter, and the entrance to the
duct is flared in order to reduce inlet losses. The
installation of the duct is judged as creating no
significant effect on the performance of the fan in this
example.
such as this example, the air may be discharging
from the fan into a region in which the atmospheric
pressure is somewhat different from that to which all
other pressure measurements are referred. When
this possibility exists, it is essential that the static
pressure in the region of the discharging air be
measured, referred to the same atmospheric
pressure as used in all other pressure
measurements. In this example, Ps2 was measured,
referred to the same atmospheric pressure as in the
static pressure measurements made at Plane 3.
4. Measure the dry-bulb and wet-bulb temperatures
at the velocity traverse plane. Determine pb for the
general vicinity of the fan. These measurements are
used to determine densities at the planes of interest.
5. Measure the fan speed and the motor amps and
volts. Record all pertinent motor nameplate data. For
the horsepower rating of the motor in this example, it
is recommended that the fan power input be
determined by using the measured watts input to the
motor and motor performance data obtained from the
motor manufacturer.
6. To calculate the Fan Static Pressure:
3. Ps2, the static pressure at the outlet of the fan, is
zero gauge pressure, referred to the atmospheric
pressure in the region of the fan outlet. In situations
236 | Field Performance Measurement
Ps = Ps2 - Ps1 - Pv1
= Ps2 - (Ps1 + Pv1)
Where:
FLOW RATES
Ps1 + Pv1 = Ps3 + Pv3
V3 = 1096 (Pv3/ρ3)0.5
= 1096 (0.025/0.0715)0.5
= 648 fpm
7. In order to compare the test results to the quoted
fan curve drawn for operation at 1725 rpm and 0.075
lbm/ft3 density, it is necessary to convert the results
to the specified conditions. The basis for the
calculations is described in Section 14.
OBSERVATIONS
SITE MEASUREMENTS
pb = 29.65 in. Hg
td3 = 85°F
tw3 = 74°F
Ps2 = 0 in. wg
Ps3 = -0.027 in. wg
Pv3 = 0.025 in. wg
N = 1775 rpm
A3 = 5.06 ft2
Q = Q1 = Q3
= V3A3
= 648 × 5.06
= 3279 cfm
FAN POWER INPUT
At the measured power input value of 637 watts, the
data supplied by the motor manufacturer indicate
efficiency of 65% for the motor.
Hmo = (637 × 0.65)/746
= 0.56 hp
Since the fan is direct connected to the motor, there
is no drive loss, and:
MEASURED MOTOR DATA
H = Hmo
= 0.56 hp
Volts = 230, 225, 230
= 228 av
Watts = 637
FAN STATIC PRESSURE
MOTOR NAMEPLATE DATA
3/4 hp, 3 phase, 60 hertz
230 volts, 1760 rpm, 4.8 FLA
GENERAL
Fan direct connected to motor. Motor efficiency data
supplied by motor manufacturer.
Ps1 + Pv1 = Ps3 + Pv3
= -0.027 + 0.025
= -0.002 in. wg
Ps = Ps2 - (Ps1 + Pv1)
= 0 - (-0.002)
= 0.002 in. wg
This small value is attributed to the loss at the duct
inlet, and the fan is considered to be operating at free
delivery (Ps = 0).
CALCULATIONS
CONVERSION TO SPECIFIED CONDITIONS
DENSITIES
For Plane 3 conditions of:
td3 = 85°F
tw3 = 74°F
p3 = pb + (Ps3/13.6)
= 29.65 + (-0.027/13.6)
= 29.65 in. Hg
Qc = 3279 (1725/1775)
= 3187 cfm
Psc = 0 in. wg
Hc = 0.56 (1725/1775)3 (0.075/0.0715)
= 0.54 hp
Use Figure N.1 in Annex N to obtain ρ3 = 0.0715
lbm/ft3.
It is assumed that ρ1 = ρ3
Field Performance Measurement | 237
EXAMPLE 5C: FREE INLET, FREE OUTLET ROOF VENTILATOR
2
1
3
COMMENTS
1. The subject of the test in this example is the roof
ventilator assembly. Before proceeding with the test,
refer to Section 17.1 for considerations affecting the
test procedure in this type of installation.
2. Ps3, the static pressure in the vicinity of the
ventilator inlet, would normally be determined by
averaging the static pressure measurements made in
a Pitot tube traverse. But in this example, a
temporary duct was not installed and the Pitot tube
traverse could not be accomplished. In this method
for testing a nonducted fan, consider the fan static
pressure (Ps) as the differential pressure, as read on
a manometer, between the pressure measured inside
the room (Ps3) and the pressure measured outside
the room in the vicinity of the ventilator outlet (Ps2).
These pressures are measured at a sufficient
distance from the ventilator so as to be unaffected by
the velocity of the entering or leaving air.
3. Ps2 is considered to be zero gauge pressure, but
since this measurement is actually part of the
differential pressure described in paragraph 2, it is
necessary to make only one density correction; the
correction is to the differential pressure, which is the
fan static pressure.
238 | Field Performance Measurement
4. Measure the dry-bulb and wet-bulb temperatures
in the region of the inside pressure measurement.
Also, determine pb in the same vicinity.
5. Measure the fan speed and the motor amps and
volts. Record all pertinent motor nameplate data. For
the horsepower rating of the motor in this example, it
is recommended that the fan power input be
determined by using the measured watts input to the
motor and motor performance data obtained from the
motor manufacturer.
6. Airflow rates are determined from the fan
manufacturer’s certified performance ratings. Draw a
fan performance curve from these ratings converted
to operation at the test values of fan speed and
entering air density. The basis for these calculations
is described in Section 14. The fan airflow rate is then
determined by entering this curve at the test values of
fan static pressure and fan power input.
OBSERVATIONS
SITE MEASUREMENTS
pb = 29.19 in. Hg
td3 = 79°F
tw3 = 63°F
Ps2 - Ps3 = 0.13 in. wg
N = 1735 rpm
MEASURED MOTOR DATA
FAN STATIC PRESSURE
Volts = 229, 229, 232
= 230 av
Watts = 1390
The fan static pressure is considered to be the
differential static pressure.
MOTOR NAMEPLATE DATA
Ps = Ps2 - Ps3
= 0.13 in. wg
1.5 hp, 3 phase, 60 hertz
230 volts, 1740 rpm, 4.8 FLA
It is assumed that Ps1 = Ps3
GENERAL
CONVERSION OF MANUFACTURER’S
RATINGS TO OPERATING CONDITIONS
Fan direct connected to motor. Motor efficiency data
supplied by motor manufacturer.
Rating Point #1
Fan performance, at standard air density, as supplied
by fan manufacturer for 1750 rpm.
Point
CFM
Ps
HP
1)
2)
3)
8900
8520
8060
0
1/8
1/4
1.45
1.50
1.55
CALCULATIONS
DENSITIES
For Plane 3 conditions of:
td3 = 79°F
tw3 = 63°F
pb3 = pb + (Ps2 - Ps1)/13.6
= 29.19 + (0.13/13.6)
= 29.2 in. Hg
Use Figure N.1 in Annex N to obtain ρ3 = 0.0715
lbm/ft3.
It is assumed that ρ1 = ρ3.
FAN POWER INPUT
At the measured power input value of 1395 watts, the
data supplied by the motor manufacturer indicate
efficiency of 77% for the motor.
Hmo = (1390 × 0.77)/746
= 1.43 hp
Since the fan is direct connected to the motor, there
is no drive loss, and:
H
= Hmo
= 1.43 hp
FAN
Q1c = 8900 (1735/1750)
= 8824 cfm
Ps1c = 0
H1c = 1.45 (1735/1750)3 (0.0715/0.0750)
= 1.35 hp
Rating Point #2
Q2c = 8520 (1735/1750)
= 8447 cfm
Ps2c = 0.125 (1735/1750)2 (0.0715/0.0750)
= 0.117 in. wg
H2c = 1.50 (1735/1750)3 (0.0715/0.0750)
= 1.39 hp
Rating Point #3
Q3c = 8060 (1735/1750)
= 7991 cfm
Ps3c = 0.25 (1735/1750)2 (0.0715/0.0750)
= 0.234 in. wg
H3c = 1.55 (1735/1750)3 (0.0715/0.0750)
= 1.44 hp
Draw a performance curve for these operating
conditions. Enter the measured values for static
pressure and horsepower on the appropriate curves.
Ideally, these two points will coincide at the same
cfm. However, usually they will not coincide and
should be averaged to determine the fan airflow rate.
If this difference is small, such as in this example, it
is only a reflection of test inaccuracies. If, however,
these differences exceed 10%, the system should be
reanalyzed for SEFs that may have been overlooked,
or for procedural errors in the initial testing.
Field Performance Measurement | 239
Qa = 8070 cfm (based upon horsepower)
Qb = 8400 cfm (based upon static pressure)
Use:
Q = 8235 cfm (average of above).
x
x BHP
x
1.25
.30
STATIC PRESSURE IN. WG (Ps)
x
1.00
.20
x
.10
SP
x
0
7000
8000
9000
CFM(Q)
Fan Performance at 0.0715 Air Density
240 | Field Performance Measurement
BHP (H)
1.50
.40
Annex B. Pitot Static Tubes
16D
8D
0.8D
0.5D Radius
0.4D
D
3D Radius
Head shall be free from nicks and burrs.
90° ± 0.1°
All dimensions shall be within ±2%.
SECTION A-A
Static Pressure
8 holes - 0.13D, not to exceed 0.04 in.,
diameter equally spaced and free from burrs.
Hole depth shall not be less than the hole
diameter.
Note: Surface finish shall be 32 micro in. or better. The static
orifices may not exceed 0.04 in. diameter. The minimum Pitot
tube stem diameter recognized under this standard shall be
0.10 in. In no case shall the stem diameter exceed 1/30 of the
test duct diameter.
Total Pressure
PITOT-STATIC TUBE WITH SPHERICAL HEAD
All other dimensions are the same
as for spherical head pitot-static
tubes.
8D
D
X
0.2D Diameter
V
X/D
V/D
X/D
V/D
0.000
0.237
0.336
0.474
0.622
0.500
0.496
0.494
0.487
0.477
1.602
1.657
1.698
1.730
1.762
0.314
0.295
0.279
0.266
0.250
0.741
0.936
1.025
1.134
1.228
0.468
0.449
0.436
0.420
0.404
1.796
1.830
1.858
1.875
1.888
0.231
0.211
0.192
0.176
0.163
1.313
1.390
1.442
1.506
1.538
1.570
0.388
0.371
0.357
0.343
0.333
0.323
1.900
1.910
1.918
1.920
1.921
0.147
0.131
0.118
0.109
0.100
ALTERNATE PITOT-STATIC TUBE WITH ELLIPSOIDAL HEAD
Figure B.1
Field Performance Measurement | 241
Annex C. Double Reverse Tubes
AIR FLOW
TUBE ENDS MUST BE SMOOTH
AND FREE FROM BURRS
IMPACT TUBE
REVERSE TUBE
SECTION VIEW
STAINLESS STEEL
TUBING PREFERRED
APPROX. 0.375 in. OD
TOTAL PRESSURE = READING A
CORRECTED FOR MANOMETER
CALIBRATION
READING A
FLEXIBLE TUBING
ING B
READ
VELOCITY PRESSURE = READING B CORRECTED FOR
MANOMETER CALIBRATION AND
CALIBRATION FACTOR FOR THE
DOUBLE REVERSE TUBE.
Notes:
1. For use in dirty or wet gas streams.
2. The double reverse tube must be calibrated and used in the same orientation as used in its calibration
3. Also referred to as impact reverse tube, combined reverse tube, and type S tube.
Figure C.1 - Double Reverse Tube
242 | Field Performance Measurement
Annex D. Pitot-Static Tube Holder
0.312 in. DIA.
PITOT-STATIC TUBE
SPLIT BRASS BUSHING
PRESS TO FIT INTO TUBING
THERMOCOUPLE
DUCT WALL
1½ in. PIPE
HALF-COUPLING
WELDED TO DUCT
BRASS
BUSHINGS
1½ in. PIPE
NIPPLE
12 in. LONG
STAINLESS STEEL TUBING
1 in. OUTSIDE DIA. × 8 ft. LONG
SLIP FIT IN BRASS BUSHINGS
Notes:
¼ in. OUTSIDE DIA.
STAINLESS STEEL TUBING
FOR GAS SAMPLING
SPLIT BRASS
BUSHING
1. Apparatus for mounting Pitot-static tube on duct
2. For use in large ducts or high velocity gas streams
3. 1 in. diameter tube slides inside 1.5 in. pipe, which
can be unscrewed and moved to another traverse
location
4. The gas sampling tube and thermocouple may be
omitted if these data are obtained in other manners
CUT-OFF AND REBRAZE
AFTER ASSEMBLY
Figure D.1 - Pitot-Static Tube Holder (Typical)
Field Performance Measurement | 243
Annex E. Static Pressure Tap
DUCT WALL
MAXIMUM 0.125 in. DIAMETER
FOR USE IN RELATIVELY
CLEAN GASES. MAY BE
NECESSARY TO INCREASE
TO 0.312 in. DIAMETER
FOR DIRTY OR WET GASES
½ in. PIPE HALF-COUPLING
OR SIMILAR ARRANGEMENT
INSIDE SURFACE OF DUCT AND
EDGE OF HOLE ARE TO BE
SMOOTH AND FREE FROM BURRS
Figure E.1 - Static Pressure Tap
MINIMUM OF FOUR TAPS,
LOCATED 90° APART AND
NEAR THE CENTER OF
EACH WALL
STATIC PRESSURE MEASUREMENT
REQUIRED AT EACH TAP. USE
THE AVERAGE OF THE MEASUREMENTS
AS THE STATIC PRESSURE FOR THE PLANE
Figure E.2 - Locations of Static Pressure Taps
244 | Field Performance Measurement
Annex F. Pitot-Static Tube Connections
PLANE 2
PLANE 1
PLANE 4
PLANE 3
*SEF 1
Ps4
FAN STATIC PRESSURE
Ps = - Ps1 - Pv1 + SEF 1
where Ps1 = Ps4
Pv1 = Pv3
Figure F.1 - Fan with Inlet Duct Only
Ps2 = 0
PLANE 3
PLANE 5
Ps3
Ps3
P v3
*SEF 1 is due to
no duct at fan outlet
PLANE 2
PLANE 1
FAN STATIC PRESSURE
Ps = Ps2
where Ps2 = Ps5
Pt1 = 0
Ps5
P v3
Figure F.2 - Fan with Outlet Duct Only
ALTERNATE
PLANE 5
PLANE 3
Ps5
PLANE 2
PLANE 1
FAN STATIC PRESSURE
Ps = Ps2 - Ps1 - Pv1
where Ps2 = Ps5
Ps1 = Ps4
Pv1 = Pv3
PLANE 4
PLANE 3
Ps3
Ps4
P v3
Figure F.3 - Fan with Inlet Duct and Outlet Duct
Field Performance Measurement | 245
Annex G. Manometer Data
10 in. wg
1:1
SLOPE
RATIO
2 in. wg
5:1 SLOPE RATIO
0.5 in. wg
20:1 SLOPE RATIO
Figure G.1 - Manometer Data
246 | Field Performance Measurement
1 in. wg
10:1 SLOPE RATIO
PERCENT UNCERTAINTY IN VELOCITY DETERMINATION
USING PITOT-STATIC TUBE AND MANOMETER DUE TO MANOMETER SLOPE
Based on an uncertainty equivalent to an indicating column length of 0.05 in. wg in a vertical manometer (1:1 slope
ratio)
.01
.02
VELOCITY PRESSURE READING, in. wg
.04 .06 0.1
0.2
0.4 0.6 1
2
0.4
0.6
0.8
3 4
6 8 10
10.0
8.0
6.0
5.0
3.0
2.0
R
TE
ME TIO
NO RA
MA OPE
SL
1:1
1.0
0.8
0.6
0.5
2:1
10
:1
:1
0.3
5:1
0.4
20
% UNCERTAINTY IN VELOCITY DETERMINATION
4.0
0.2
0.3
1
2
3
4
6
8
10
15
STANDARD AIR VELOCITY, fpm (×1000)
Figure G.2 - Uncertainty in Velocity Determination
Field Performance Measurement | 247
Annex H. Distribution of Traverse Points
In order to obtain a representative average velocity in a duct, it is necessary to locate each traverse point
accurately. It is recommended that the number of traverse points increase with increasing duct size. The
distributions of traverse points for circular ducts, as indicated below, are based on log-linear Pitot traverse method.
X1
60º
X2
X3
X4
D
Xn
Xa = D × Ka
Where:
D is the inside diameter of the duct
Ka is the factor corresponding to the duct size and the traverse point location as indicated in the table below
NUMBER OF
INSIDE
TRAVERSE
DIAMETER POINTS IN K1
OF DUCT EACH OF 3
DIAMETERS
K2
K3
K4
K5
K6
K7
K8
K9
K10
K11
K12
K13
K14
K15
K16
LESS THAN
8 ft.
8
.021 .117 .184 .345 .655 .816 .883 .979
8 ft.
THROUGH
12 ft.
12
.014 .075 .114 .183 .241 .374 .626 .759 .817 .886 .925 .986
GREATER
THAN 12 ft.
16
.010 .055 .082 .128 .166 .225 .276 .391 .609 .724 .775 .834 .872 .918 .945 .990
Figure H.1 - Distribution of Traverse Points for Circular Ducts
248 | Field Performance Measurement
The recommended minimum number of traverse points for rectangular ducts is indicated below in Figure H.3. For
rectangular ducts with cross-sectional areas of 24 square feet and less, the recommended minimum number is 24.
For cross-sectional areas greater than 24 square feet, the minimum number of points increases as indicated in
Figure H.3. The points are to be located in the centers of equal areas with the areas as nearly square as practical
(see Figure H.2). If the flow conditions at the traverse plane are less than satisfactory, the accuracy of the
determination of flow rate may be improved by using more than the recommended minimum number of points.
Fewer points may be used if the flow is very uniform; however, the maximum area covered per point should not
exceed 3 square feet.
Y
Y
2
X
2
X
Figure H.2 - Distribution of Traverse Points for Rectangular Duct
NUMBER OF TRAVERSE POINTS
100
90
80
70
60
50
40
30
25
20
15
10
10
15
20
25 30
40
50 60 70 80 100
150
200 250 300
DUCT CROSS-SECTIONAL AREA, ft2
Figure H.3 - Recommended Minimum Number
of Traverse Points for Rectangular Ducts
Field Performance Measurement | 249
Annex J. Instrumentation Characteristics
Table J.1 - Temperature Measurement
Approximate
Range
F
Precision
F
Limitations
Temp of gases and liquids by contact
-38/575
Less than
0.1 to 10
In gases, accuracy affected by radiation
”
”
-100/100
-200/70
”
-38/1000
-459/1000
No. Measurement Means
1. Glass-stem thermometers
Mercury-glass thermometer
Application
Alcohol-glass thermometer
Pentane-glass thermometers
Jena or quartz mercury nitrogen
thermometers
2. Gas thermometer
3. Resistance thermometers
Platinum-resistance thermometer
Primary standard
”
”
”
”
”
”
Less than
0.01
Requires considerable skill to use
High cost; accuracy
affected by radiation
in gases
Accuracy affected by
radiation in gases
Precision; remote readings; temp of
fluids or solids by contact
-320/1800
Less than
0.02 to 5
Remote readings; temp by contact
-150/300
0.3
Up to 600
0.1
Standard for thermocouples
500/3000
0.1 to 5
General testing of high temp; remote
rapid readings by direct contact
”
Same as above, especially suited for
low temp
Up to 2200
0.1 to 15
Up to 1500
Up to 700
0.1 to 15
0.1 to 15
5. Beckman thermometers
(metastatic)
For differential temp in same applications as in glass stem thermometer
9 diff
0.018
6. Bimetallic thermometers
For approx temp
0/1000
1, usually
much more
7. Pressure-bulb thermometers
Gas-filled bulb
Remote-testing
-100/1000
2
Caution must be exercised so that installation is correct
”
”
For intensity of narrow spectra band
of high temp radiation (remote)
20/500
-50/2100
2
2
15
”
”
9. Radiation pyrometers
For intensity of total high temp radiation (remote)
Any range
10. Seger cones (fusion pyrometers)
Approx temp (within temp source)
1000/3600
50
11. Indicating crayons
Approx temp (in surface)
125/900
12. Melting and boiling points of
materials
Standards
All except extremely high
temp
±1%
Extremely
precise
Nickel-resistance thermometer
Thermistors
4. Thermocouples
Pt-Pt-Rh thermocouple
Chromel-alumel thermocouple
Iron-constantain thermocouple
Copper-constantan thermocouple
Chromel-constantan thermocouple
Vapor-filled bulb
Liquid-filled bulb
8. Optical pyrometers
”
1500 upward
Reprinted by permission from ASHRAE Handbook - 1989 Fundamentals
250 | Field Performance Measurement
High cost; also, requires expensive
measuring device
Less accurate than
above
Subject to oxidation
Must be set for temp
to be measured
Time lag; unsuitable
for remote use; unreliable
For laboratory use
only
Table J.2 - Differential Pressure Measurement
No.
Range
Precision
Limitations
1. Micromanometer
Measurement Means
Very low press. diff.
Application
0 to 6 in. H20
0.005 to
0.001 in. H20
Not readily portable; not easy to
use with pulsating pressure
2. Draft gauges
Moderately low press. diff.
0 to 10 in. H20
0.005 to
0.05 in. H20
Must be leveled carefully
3. Manometer
Medium press diff.
0 to 100 in. H20
or Hg
0.05 in.
Where used with liquid must be
compensated for liquid density
4. Swinging-vane-type gauge
Moderately low press. diff.
0 to 0.5 in. H20
0 to 20 in. H20
5%
Generally usable to atmospheric
pressure only
5. Bourdon-tube type
Medium to high press. diff.,
usually to atmosphere
Any
0.05 to 5%
Subject to damage due to over
press-shock or pulsation
6. Pressure transducersstrain gauge, capacity, potentiometer, crystal, magnet
Remote reading, responds
to rapid changes of pressure
0.05 to 50,000
psi
0.1 to 0.5%
Requires electronic amplifier and
readout device
Table J.3 - Velocity Measurement
No.
Application
Range
Precision
Limitations
1. Smoke puff or airborne
solid tracer
Measurement Means
Low air velocities in rooms;
highly directional
5 to 50
10 to 20%
Awkward to use but valuable in
tracing air movement
2. Deflecting-vane
anemometer
Air velocities in rooms, at
outlets, etc; directional
30 to 24,000
5%
Not well suited for duct readings;
needs periodic check calibration
3. Revolving-vane
anemometer
Moderate air velocities in
ducts and rooms; somewhat directional
100 to 3000
5 to 20%
Extremely subject to error with
variations in velocities with space
or time; easily damaged; needs
periodic calibration
4. Pitot tube
Std instrument for measurement of duct velocities
180 to 10,000
with micromanometer
600 to 10,000 with
draft gauges; 10,000
up with manometer
1 to 5%
Accuracy falls off at low end of
range
5. Impact tube and sidewall or other static tap
High velocities, small
tubes and where air direction may be variable
120 to 10,000
with micromanometer;
600 to 10,000 with
draft gauges; 10,000 up
with manometer
1 to 5%
Accuracy depends upon constancy
of static pressure across stream
section
6. Heated thermocouple
anemometer
Air velocities in ducts,
velocity distributions
10 to 2000
3 to 20%
Accuracy of some types not good
at lower end of range; steady
state measurements only
7. Hot-wire anemometer
(a) Low air velocities; directional and nondirectional available
1 to 1000
1 to 20%
Requires accurate calibration at
frequent intervals; complex,
costly
(b) High air velocities
up to 60,000
1 to 20%
(c) Transient velocity and
turbulence
Reprinted by permission from ASHRAE Handbook - 1989 Fundamentals
Field Performance Measurement | 251
Annex K. Phase Current Method for
Estimating the Power Output of Three
Phase Fan Motors
Use Equation A to estimate the Hmo for motors of 5
horsepower and greater, operating at 90% or more of
FLA. The uncertainties will be less than 5%.
The power output of three phase motors can be
estimated based on the relationship of motor current
and motor power output. Two equations can be used
in estimating the motor power output. The equations
are as follows:
Use the average of Equation A and Equation B to
estimate the Hmo for all motors operating at less than
90% of FLA and for 3 horsepower and smaller motors
operating above 90% of FLA. An estimated Hmo less
than 50% of NPH can contain 15% uncertainties or
greater.
Equation A:
⎛ Measured amps ⎞ ⎛ Measured volts ⎞
Hmo = NPH ⎜
⎟⎜
⎟
FLA
NPV
⎝
⎠⎝
⎠
Where:
Hmo = motor power output
NPH = nameplate horsepower
FLA = full load amps
NPV = nameplate volts
measured volts = average of the measured phase
volts
measured amps = average of the measured phase
amps
Equation B:
⎛ Measured amps - NLA ⎞ ⎛ Measured volts ⎞
Hmo = NPH ⎜
⎟⎜
⎟
FLA - NLA
NPV
⎝
⎠⎝
⎠
Where:
NLA = average of the measured phase values of no
load amps
NPH = nameplate horsepower
FLA = full load amps
NPV = nameplate volts
NLA can usually be obtained with the motor operating
and the motor shaft coupling or belt drive
disconnected. In the case where the fan impeller is
mounted directly on the motor shaft, it will be
necessary to remove the impeller in order to obtain
NLA measurements.
252 | Field Performance Measurement
Figure K.1 represents the relationship of motor
current and motor power output. The “dashed” lines
between 0% NPH and 100% NPH for motor sizes
shown represents Equation B. The solid lines
between these same end points for the motor sizes
shown represent the general shape of typical motor
calibration amp/load curves. The solid line from
100% NPH and 100% FLA to 0% NPH and 0% FLA
represents Equation A. These curves indicate that if
you average the results of Equation A and Equation
B for a specific measured amp draw, that your results
approach the typical calibration curve. It also points
out that the uncertainties are low if just Equation A is
used above 90% FLA, especially in the larger integral
motor horsepowers.
Many fractional horsepower and small integral
horsepower motors do not have a significant change
in current from no load to full load. The actual ampsload characteristics for motors of the same
horsepower rating can vary greatly from motor
manufacturer to motor manufacturer. No load
amperage (NLA) varies significantly for the same size
motor between manufacturers. In addition, various
motor design requirements result in different ampload characteristics even though the horsepower
ratings of the motors are the same. These are some
of the reasons that Figure K.1 cannot be used to
determine the motor output directly. The chart is only
intended to indicate the accuracy and suitability of
using the above equations for estimating motor
power output.
GENERALIZED CURVES ILLUSTRATING THE RELATIONSHIP OF
HORSEPOWER TO AMPS FOR THREE PHASE MOTORS
Do not use for determining actual motor horsepower
DOTTED LINES PER EQUATION B: Hmo ∝ MEASURED AMPS - NLA/FLA - NLA
100
90
RATED
HORSEPOWER
1
2
80
70
3
60
5
50
10
40
400
30
2500
20
10
0
0
10
20
30
40
50
60
70
80
90
100
% NAMEPLATE HORSEPOWER
PER EQUATION A: Hmo ∝
MEASURED AMPS
FLA
CAUTION: THIS CHART IS REPRESENTATIVE ONLY! SINCE THE AMP-LOAD CHARACTERISTICS OF THE
SAME SIZE MOTOR WILL VARY BETWEEN THE VARIOUS MOTOR MANUFACTURERS, IT CANNOT BE USED
TO DETERMINE THE HORSEPOWER OUTPUT OF A MOTOR. USE THE EQUATIONS AS DIRECTED ON THE
PREVIOUS PAGE.
Field Performance Measurement | 253
Annex L. Estimated Belt Drive Loss
Drive loss is defined as follows:
Percent drive loss equals power to driving sheave
minus power from driven sheaves times 100, divided
by power to driving sheave.
There are several things which can affect belt drive
efficiencies. Some of these are:
1) Over-designed drives. This was considered good
practice at one time because the drive would last
longer. It will still last longer but it is more
inefficient.
2) Multiple belts on subminimum diameter sheaves
are less efficient than fewer belts on larger
diameter sheaves. Both the National Electric
Motor
Association
and
the
Rubber
Manufacturer’s Association publish data dealing
with minimum recommended sheave diameters.
As these minimum sheave diameters are
approached, the drive loss becomes greater.
3) A larger belt section than required will increase
the drive loss.
4) A badly undertensioned drive will increase the
drive loss.
5) Misaligned drives will increase the drive loss.
Drive loss is manifested as heat in belt drives. Under
ambient conditions of less than 100°F, well designed
drives that operate efficiently will be warm to the
touch immediately after being shut down. If the drive
is uncomfortable to the touch (approximately 140°F
or more), then the drive loss is high. Obviously poorly
tensioned and misaligned drives should be corrected
before estimating brake horsepowers and drive
losses.
254 | Field Performance Measurement
100
80
60
DRIVE LOSS, % MOTOR POWER OUTPUT*
40
30
RANGE OF DRIVE LOSSES FOR STANDARD BELTS
20
15
10
8
6
4
3
2
1.5
1
0.3 0.4 0.6 0.8 1
2
3
4
6
8 10
20
30 40
60 80 100
200 300 400
600
MOTOR POWER OUTPUT, hp
HIGHER BELT SPEEDS TEND TO HAVE HIGHER LOSSES
THAN LOWER BELT SPEEDS AT THE SAME HORSEPOWER
*Drive losses are based on the conventional V-belt, which has been the “work horse” of the drive industry for
several decades.
EXAMPLE
• Motor power output, Hmo, is determined to be 13.3 hp
• The belts are the standard type and just warm to the touch immediately after shutdown
• From chart, drive loss = 5.1%
• Drive loss, HL
= 0.051 × 13.3
= 0.7 hp
• Fan power input, H
= 13.3 - 0.7
= 12.6 hp
Figure L.1 - Estimated Belt Drive Loss
Field Performance Measurement | 255
Annex M. Density Determinations
Since:
M.1 General
Ps1 = 0
p1 = pb
= 28.60 in. Hg
This annex contains examples illlustrating the
procedures for determining densities. Determinations
of densities are shown for air and for gases other
than air.
M.2 Determination of the density of air,
general case
Determine air density by using the Psychrometric
Density Chart, shown in Figure N.1 in Annex N, the
Psychrometric Density Table, shown in Annex N, or a
calculation procedure which makes use of perfect
gas relationships and the modified Apjohn equation
for partial vapor pressure. Examples of the use of
these procedures are included in this section. Each
of the procedures requires knowledge of the
pressure, dry-bulb temperature and wet-bulb
temperature of the air.
The Psychrometric Density Chart and the
Psychrometric Density Table are limited to the
temperatures and pressures normally encountered in
fan applications.
Limit the use of the calculation procedure that is
based on perfect gas relationships and illustrated in
Example M2.3, to instances in which the dry-bulb
temperature is 180°F or less. Accurate wet-bulb
temperature measurements are difficult to obtain
when the dry-bulb temperature exceeds 180°F.
When the dry-bulb temperature exceeds 180°F, it
may be necessary to rely on site personnel for the
water vapor content of the air. Alternately,
commercially available instrumentation for dew point
determination may be used. For the procedure
required to determine density based on the data
provided in either of the above cases, refer to
Psychrometric Tables and Charts by Zimmerman and
Lavine.1
EXAMPLE M2.1
The conditions that exist at the inlet of a fan that is
not ducted on the inlet side are:
td1 = 78°F
tw1 = 62°F
The wet-bulb depression is:
td1 - tw1 = 78 - 62
= 16°F
For wet-bulb depression of 16°F, dry-bulb
temperature of 78°F and absolute pressure of 28.60
in. Hg, obtain ρ1 = 0.0701 lbm/ft3 by using the
Psychrometric Density Chart in Figure N.1 in Annex N.
EXAMPLE M2.2
The conditions at a fan inlet, located at an elevation
of 1000 ft above sea level are:
Ps1 = -3.45 in. wg
td1 = 85°F
tw1 = 75°F
Barometric pressure, obtained from a nearby airport,
is 29.82 in. Hg at sea level.
Using the data in Figure N.3 in Annex N, the
barometric pressure at 1000 ft above sea level is:
pb = 29.82 × 0.964
= 28.75 in. Hg
The absolute pressure at the fan inlet is:
p1 = pb + (Ps1/13.6)
= 28.75 + (-3.45/13.6)
= 28.50 in. Hg
The wet-bulb depression is:
td1 - tw1 = 85 - 75
= 10°F
For dry-bulb temperature of 85°F, absolute pressure
of 28.50 in. Hg and wet-bulb depression of 10°F, use
the Psychrometric Density Table in Figures N.5 in
Annex N to obtain:
ρ1 = 0.06829 + 10 × 0.000041
= 0.0687 lbm/ft3
1. O. T. Zimmerman and I. Lavine, Psychrometric Tables and Charts, 2nd ed. (Dover, N.H.: Industrial Research Service Inc., 1964)
256 | Field Performance Measurement
Example M2.3
EXAMPLE M3.1
The conditions at a fan inlet are:
Dry air is entering a fan inlet, located at an elevation
of 1000 ft above sea level. The pressure and
temperature at the inlet are:
Ps1 = -8.75 in. wg
td1 = 146°F
tw1 = 93°F
The barometric pressure, pb, measured for the
atmosphere to which Ps1 is referred, is 28.15 in. Hg.
The absolute pressure at the fan inlet is:
p1 = pb + (Ps1 /13.6)
= 28.15 + (-8.75/13.6)
= 27.51 in. Hg
Use Figure N.2 in Annex N to obtain saturated vapor
pressure, pe, of 1.562 in. Hg for the wet-bulb
temperature of 93°F.
Use the modified Apjohn equation for partial vapor
pressure, pp, to obtain:
pp = pe - p1 (td1 - tw1)/2700
= 1.562 - 27.51 (146 - 93)/2700
= 1.022 in. Hg
ρ1 is calculated by using perfect gas relationships:
ρ1 =
=
1.3257 ( p1 − 0.378 pp )
( td1 + 460 )
1.3257 ( 27.51 − 0.378 × 1.022 )
(146 + 460 )
= 0.0593 lbm/ft 3
M.3 Determination of the density of air,
special cases
The procedures for the determination of the density
of air that are described in Section M.2 are valid for
dry air, air that is saturated with water vapor and air
that is partially saturated with water vapor. This
section contains alternate procedures for cases in
which it is known that the air is either dry or saturated.
Knowledge that the air is either dry or saturated
eliminates the usual requirement of the wet-bulb
temperature determination; however, it should be
noted that an incorrect assumption of either of these
conditions can result in a significant uncertainty in the
density determination.
Ps1 = -15 in. wg
td1 = 95°F
Barometric pressure, obtained from a nearby airport,
is 29.24 in. Hg at sea level.
Using the data in Figure N.3 in Annex N, the
barometric pressure at 1000 ft above seal level is:
pb = 29.24 × 0.964
= 28.19 in. Hg
The absolute pressure at the fan inlet is:
p1 = pb + (Ps1/13.6)
= 28.19 + (-15/13.6)
= 27.09 in. Hg
Dry air at 29.92 in. Hg and 70°F has a density of
0.075 lbm/ft3.
Consider the density of air to be directly proportional
to absolute pressure and inversely proportional to
absolute temperature. The density of the air at the fan
inlet is calculated as follows:
ρ1 = 0.075 (p1/29.92) [(70 + 460)/(td1 + 460)]
= 0.075 (27.09/29.92) [530/(95 + 460)]
= 0.0648 lbm/ft3
EXAMPLE M3.2
Saturated air is enterting a fan inlet, located at an
elevation of 1500 ft above sea level. The pressure
and temperature at the inlet are:
Ps1 = - 6.75 in. wg
td1 = 103°F
Barometric pressure, obtained from a nearby airport,
is 29.66 in. Hg at sea level.
Using the data in Figure N.3 in Annex N, the
barometric pressure at 1500 ft above sea level is:
pb = 29.66 × 0.947
= 28.09 in. Hg
The absolute pressure at the fan inlet is:
p1 = pb + (Ps1/13.6)
Field Performance Measurement | 257
= 28.09 + (-6.75/13.6)
= 27.59 in. Hg
Refer to Figure N.4 in Annex N to obtain saturated air
density of 0.06868 at 103°F and 29.92 in. Hg.
EXAMPLE M4.1
A gas is entering a fan inlet located at an elevation of
2000 ft above sea level. The pressure and
temperature at the inlet are:
Assuming the density of saturated air to be directly
proportional to absolute pressure, the density at the
fan inlet is calculated as follows:
Ps1 = - 22 in. wg
td1 = 230°F
ρ1 = 0.06868 (p1/29.92)
= 0.06868 (27.59/29.92)
= 0.0633 lbm/ft3
Barometric pressure, obtained from a nearby airport,
is 29.92 in. Hg at sea level. The composition of the
gas is 5.5% CO2, 1% CO, 15% O2, 1% H2, and 77.5%
N2, by volume.
Assuming the density of saturated air to be directly
proportional to absolute pressure is an
approximation. The uncertainty in the density
determination as a result of this approximation
increases with increasing temperature and increases
with increasing variation between the actual absolute
pressure and 29.92 in. Hg, which is the stated
pressure for the data in Figure N.4. The uncertainty
will be approximately 1% or less under the following
conditions:
•
•
•
At 120°F and at an absolute pressure within 20%
of 29.92 in. Hg
At 150°F and at an absolute pressure within 10%
of 29.92 in. Hg
At 180°F and at an absolute pressure within 4% of
29.92 in. Hg
The apparent molecular weight of the gas is
determined as follows:
Volume
Molecular
Component Fraction × Weight = lb/mole
CO2
CO
O2
H2
N2
0.055
0.01
0.15
0.01
0.775
44
28
32
2
28
1.00
2.42
0.28
4.80
0.02
21.70
29.22
Apparent molecular weight = (29.22/1.00)
= 29.22
M.4 DETERMINATION OF THE DENSITY OF A
GAS OTHER THAN AIR
The density of the gas at 70°F and 29.92 in. Hg is
calculated as follows:
The determination of the density of a gas other than
air may require the use of complex equipment.
Unless specifically qualified, an expert should be
consulted for the proper use of the equipment. If the
gas is a complex mixture of various consitutuents, as
found in certain industrial processes, it is suggested
that the company chemist be consulted for the gas
analysis. Particular care should be used if the gas is
toxic, corrosive or explosive; and in these cases,
consideration should be given to substituting air for
the test.
Apparent molecular weight 29.22
=
386.7
386.7
The first two examples in this section illustrate gas
density determinations based on analyses that
provide the relative amounts of the gas constituents.
Typical flue gas density data, which is provided in
Figure N.6 in Annex N, is illustrated in Example M4.3.
Since the actual density may be significantly different
from the density determined by using typical data, it
is recommended that the typical data be used only in
the even that more specific information is not
available.
258 | Field Performance Measurement
= 0.0756 lbm/ft 3
Using the data in Figure N.3 in Annex N, the
barometric pressure at 2000 ft above sea level is:
pb = 29.92 × 0.930
= 27.83 in. Hg
The absolute pressure at the fan inlet is:
p1 = pb + (Ps1/13.6)
= 27.83 + (-22/13.6)
= 26.21 in. Hg
Consider the density of the gas to be directly
proportional to absolute pressure and inversely
proportional to absolute temperature. The density of
the gas at the fan inlet is calculated as follows:
ρ1 = 0.0756 (p1/29.92)[(70 + 460)/(td1 + 460)]
= 0.0756 (26.21/29.92) [530/(230 + 460)]
= 0.0509 lbm/ft3
EXAMPLE M4.2
The conditions that exist at the inlet of a fan are Ps1 =
-19.5 in. wg and td1 = 240°F. The barometric pressure,
pb, measured for the atmospheric to which Ps1 is
referred is 29.35 in. Hg. The composition of the gas
is 5.5% CO2, 1% CO, 15% O2, 1% H2, and 77.5% N2
by weight.
The apparent molecular weight of the gas is
determined as follows:
Volume
Molecular
Component Fraction × Weight = lb/mole
CO2
CO
O2
H2
N2
0.055
0.01
0.15
0.01
0.775
44
28
32
2
28
1.00
0.00125
0.00036
0.0047
0.005
0.0277
0.0390
Apparent molecular weight = 1/0.0390
= 25.6
The density of the gas at 70°F and 29.92 in. Hg is
calculated as follows:
EXAMPLE M4.3
Flue gas is flowing at Plane 3, the Pitot traverse
measurement plane. The flue gas is the result of
using natural gas as the fuel. The conditions that
exsit at Plane 3 are:
Ps3 = 5.74 in. wg
td3 = 680°F
The barometric pressure, pb, measured for the
atmosphere to which Ps3 is referred is 28.85 in. Hg.
The absolute pressure at Plane 3 is:
p3 = pb + (Ps3/13.6)
= 28.85 + (5.74/13.6)
= 29.27 in. Hg
Refer to Figure N.6 in Annex N to obtain typical flue
gas density when natural gas is used as the fuel of
0.0725 lbm/ft3 at 70°F and 29.92 in. Hg.
Consider the density of the flue gas to be directly
proportional to absolute pressure and inversely
proportional to absolute temperature. The density of
the gas at Plane 3 is calculated as follows:
ρ1 = 0.0725 (p3/29.92)[(70 + 460)/(td3 + 460)]
= 0.0725 (29.27/29.92) [530/(680 + 460)]
= 0.0330 lbm/ft3
Apparent molecular weight 25.6
=
386.7
386.7
= 0.0662 lbm/ft 3
The absolute pressure at the fan inlet is:
p1 = pb + (Ps1/13.6)
= 29.35 + (-19.5/13.6)
= 27.92 in. Hg
Consider the density of the gas to be directly
proportional to absolute pressure and inversely
proportional to absolute temperature. The density of
the gas at the fan inlet is calculated as follows:
ρ1 = 0.0662 (p1/29.92)[(70 + 460)/(td1 + 460)]
= 0.0662 (27.92/29.92) [530/(240 + 460)]
= 0.0468 lbm/ft3
Field Performance Measurement | 259
76
74
72
70
68
66
64
62
60
58
56
54
52
50
48
46
44
42
0
DRY-BULB TEMPERATURE, °F
0.080
0.079
78
2
80
0.078
4
82
0.077
6
84
0.076
8
86
0.075
10
0.070
0.069
0.068
0.067
0.066
0.074
88
Wet-bulb depression = 4°F; proceed horizontally to 54°F dry-bulb temperature;
read vertically to 29.9 in. Hg; read horizontally to the density -- ρ = 0.0769 lbm/ft3.
12
90
td = 54°F; tw = 50°F; pb = 29.9 in. Hg
Given:
Solution:
•
•
0.065
0.064
0.073
92
4.
14
94
Read vertically to the absolute pressure.
Then read horizontally to the density.
3.
0.063
0.062
0.072
96
Proceed horizontally to the appropriate dry-bulb
temperature.
2.
Example
Calculate wet-bulb depression. Enter chart at the left.
1.
0.061
0.060
16
g
0.071
98
28.0
28.2
28.4
28.6
28.8
29.0
29.2
29.4
29.6
29.8
30.0
in. H
18
20
22
24
26
28
30
32
34
36
38
OLU
ABS
40
TE
PRE
SS
URE
260 | Field Performance Measurement
AIR DENSITY, lbm/ft3
WET-BULB DEPRESSION, °F
Figure N.1 - Psychrometric Density Chart
Field Performance Measurement | 261
tw
°F
pe
in. Hg
tw
°F
pe
in. Hg
tw
°F
pe
in. Hg
tw
°F
pe
in. Hg
tw
°F
pe
in. Hg
30
31
32
33
34
.1646
.1724
.1805
.1879
.1956
60
61
62
63
64
.5219
.5408
.5603
.5804
.6011
90
91
92
93
94
1.423
1.468
1.515
1.562
1.611
120
121
122
123
124
3.451
3.548
3.647
3.749
3.853
150
151
152
153
154
7.580
7.770
7.963
8.161
8.362
35
36
37
38
39
.2036
.2118
.2204
.2292
.2384
65
66
67
68
69
.6225
.6445
.6667
.6906
.7148
95
96
97
98
99
1.662
1.714
1.767
1.821
1.877
125
126
127
128
129
3.960
4.069
4.180
4.295
4.412
155
156
157
158
159
8.569
8.779
8.994
9.213
9.437
40
41
42
43
44
.2478
.2576
.2678
.2783
.2892
70
71
72
73
74
.7397
.7653
.7917
.8188
.8468
100
101
102
103
104
1.935
1.994
2.054
2.117
2.180
130
131
132
133
134
4.531
4.654
4.779
4.908
5.038
160
161
162
163
164
9.665
9.898
10.14
10.38
10.63
45
46
47
48
49
.3004
.3121
.3241
.3365
.3494
75
76
77
78
79
.8757
.9053
.9359
.9673
.9997
105
106
107
108
109
2.246
2.313
2.381
2.452
2.525
135
136
137
138
139
5.173
5.310
5.450
5.593
5.740
165
166
167
168
169
10.88
11.13
11.40
11.66
11.94
50
51
52
53
54
.3626
.3764
.3905
.4052
.4203
80
81
82
83
84
1.033
1.067
1.103
1.139
1.176
110
111
112
113
114
2.599
2.675
2.753
2.833
2.915
140
141
142
143
144
5.889
6.043
6.199
6.359
6.522
170
171
172
173
174
12.21
12.50
12.79
13.08
13.38
55
56
57
58
59
.4359
.4520
.4687
.4859
.5036
85
86
87
88
89
1.214
1.254
1.294
1.336
1.379
115
116
117
118
119
2.999
3.085
3.173
3.263
3.356
145
146
147
148
149
6.689
6.860
7.034
7.212
7.394
175
176
177
178
179
180
13.69
14.00
14.32
14.64
14.94
15.31
Adapted from ASHRAE Handbook - 1989 Fandamentals
Figure N.2 - Thermodynamic Properties of Water at Absolute Vapor Pressures, Inches of Mercury
262 | Field Performance Measurement
ALTITUDE
ft.
SPECIFIC
GRAVITY
PRESSURE
in. Hg
ALTITUDE
ft.
SPECIFIC
GRAVITY
PRESSURE
in. Hg
0
100
200
300
400
1.00
0.996
0.993
0.989
0.986
29.92
29.81
29.70
29.60
29.49
3000
3200
3400
3600
3800
0.896
0.890
0.833
0.877
0.870
26.82
26.62
26.42
26.23
26.03
500
600
700
800
900
0.982
0.979
0.975
0.971
0.968
29.38
29.28
29.17
29.07
28.96
4000
4200
4400
4600
4800
0.864
0.857
0.851
0.845
0.838
25.84
25.65
25.46
25.27
25.08
1000
1100
1200
1300
1400
0.964
0.961
0.957
0.954
0.950
28.86
28.75
28.65
28.54
28.44
5000
5200
5400
5600
5800
0.832
0.826
0.820
0.814
0.807
24.90
24.71
24.52
24.34
24.16
1500
1600
1700
1800
1900
0.947
0.944
0.940
0.937
0.933
28.33
28.23
28.13
28.02
27.92
6000
6500
7000
7500
8000
0.801
0.786
0.772
0.757
0.743
23.98
23.53
23.09
22.65
22.22
2000
2100
2200
2300
2400
0.930
0.926
0.923
0.920
0.916
27.82
27.72
27.62
27.52
27.42
8500
9000
9500
10000
15000
0.729
0.715
0.701
0.688
0.564
21.80
21.39
20.98
20.58
16.89
2500
2600
2700
2800
2900
0.913
0.909
0.906
0.903
0.899
27.32
27.21
27.11
27.01
26.91
20000
25000
30000
35000
40000
0.460
0.371
0.297
0.235
0.185
13.75
11.10
8.89
7.04
5.54
Note: Specific gravity of standard air at sea level and 29.92 in. Hg = 1.00
Figure N.3 - Relative Specific Gravity of Air at Various Altitudes1
1. Robert Jorgensen, ed., Fan Engineering, 7th ed. (Buffalo, NY, Buffalo Forge Co., 1970) p.8 - Reprinted by Permission
Field Performance Measurement | 263
PROPERTIES OF SATURATED AIR2
Temp
°F
WEIGHT IN A
CUBIC FOOT
OF MIXTURE
WEIGHT OF
THE VAPOR
VOLUME
ft3/lb
OF
TOTAL
lb/lb
lb/lb
DRY AIR VAPOR
WEIGHT DRY AIR
OF
OF
lb
lb
lb
DRY AIR MIXTURE
Temp
°F
WEIGHT IN A
CUBIC FOOT
OF MIXTURE
VOLUME
ft3/lb
WEIGHT OF
THE VAPOR
OF
TOTAL
lb/lb
lb/lb
DRY AIR VAPOR
WEIGHT DRY AIR
OF
OF
lb
lb
lb
DRY AIR MIXTURE
-25
-20
-15
-10
-5
.09134
.09025
.08922
.08820
.08723
.000018
.000024
.000031
.000041
.000053
.09136
.09027
.08925
.08824
.08728
10.95
11.07
11.21
11.34
11.46
.00020
.00027
.00035
.00046
.00061
.00020
.00027
.00035
.00046
.00061
46
47
48
49
50
.07768
.00750
.07731
.07714
.07694
.000509
.000527
.000545
.000567
.000587
.07819
.07803
.07785
.07771
.07753
12.87
12.90
12.93
12.96
12.99
.00655
.00680
.00705
.00734
.00762
.00651
.00675
.00700
.00728
.00756
0
5
10
15
20
.08625
.08529
.08434
.08340
.08247
.000068
.000087
.000110
.000140
.000176
.08632
.08538
.08445
.08354
.08264
11.59
11.72
11.85
11.99
12.12
.00080
.00102
.00130
.00168
.00213
.00080
.00102
.00130
.00168
.00213
51
52
53
54
55
.07676
.07657
.07637
.07620
.07600
.000608
.000632
.000651
.000675
.000700
.07737
.07720
.07702
.07687
.07670
13.02
13.06
13.09
13.12
13.15
.00792
.00823
.00854
.00884
.00921
.00786
.00819
.00845
.00877
.00913
21
22
23
24
25
.08230
.08210
.08193
.08173
.08156
.000185
.000193
.000202
.000213
.000222
.08248
.08229
.08213
.08194
.08178
12.15
12.18
12.20
12.23
12.26
.00225
.00235
.00246
.00260
.00272
.00224
.00234
.00245
.00259
.00271
56
57
58
59
60
.07582
.07562
.07544
.07524
.07506
.000723
.000749
.000775
.000801
.000829
.07654
.07637
.07622
.07604
.07589
13.19
13.22
13.25
13.29
13.32
.00952
.00989
.01026
.01063
.01103
.00943
.00980
.01016
.01052
.01091
26
27
28
29
30
.08136
.08117
.08099
.08083
.08063
.000233
.000243
.000254
.000264
.000277
.08159
.08141
.08124
.08109
.08090
12.29
12.32
12.34
12.37
12.40
.00285
.00300
.00314
.00328
.00345
.00284
.00299
.00313
.00327
.00344
61
62
63
64
65
.07486
.07468
.07447
.07429
.07408
.000857
.000886
.000916
.000947
.000979
.07572
.07557
.07539
.07524
.07506
13.35
13.39
13.42
13.46
13.49
.01143
.01185
.01229
.01273
.01320
.01130
.01171
.01214
.01257
.01303
31
32
33
34
35
.08043
.08025
.08006
.07989
.07970
.000290
.000303
.000315
.000327
.000339
.08072
.08055
.08038
.08022
.08004
12.43
12.46
12.49
12.51
12.54
.00362
.00378
.00393
.00409
.00426
.00361
.00376
.00392
.00408
.00425
66
67
68
69
70
.07390
.07369
.07350
.07330
.07310
.001012
.001045
.001080
.001115
.001152
.07491
.07473
.07458
.07441
.07425
13.53
13.57
13.60
13.64
13.68
.01368
.01417
.01468
.01520
.01576
.01349
.01397
.01447
.01497
.01551
36
37
38
39
40
.07952
.07933
.07916
.07897
.07880
.000353
.000364
.000380
.000394
.000409
.07987
.07969
.07954
.07936
.07921
12.57
12.60
12.63
12.66
12.69
.00444
.00460
.00480
.00499
.00519
.00442
.00458
.00478
.00496
.00516
71
72
73
74
75
.07290
.07270
.07250
.07229
.07208
.001189
.001229
.001268
.001310
.001352
.07409
.07393
.07377
.07360
.07343
13.71
13.75
13.79
13.83
13.87
.01630
.01691
.01748
.01812
.01876
.01604
.01662
.01717
.01780
.01841
41
42
43
44
45
.07860
.07843
.07825
.07805
.07788
.000425
.000440
.000456
.000473
.000491
.07902
.07887
.07871
.07852
.07837
12.72
12.75
12.78
12.81
12.84
.00541
.00561
.00583
.00606
.00630
.00538
.00558
.00579
.00602
.00626
76
77
78
79
80
.07188
.07166
.07144
.07124
.07104
.001395
.001439
.001485
.001532
.001579
.07328
.07310
.07293
.07277
.07262
13.91
13.95
13.99
14.03
14.08
.01941
.02008
.02079
.02150
.0223
.01904
.01968
.02036
.02106
.02174
Figure N.4 - Weights of Air, Water Vapor, and Saturated Mixture of Air and
Water Vapor at Different Temperatures and 29.92 in. Hg
2. Jorgensen, op. cit., pp 15-17
264 | Field Performance Measurement
Reprinted by Permission
PROPERTIES OF SATURATED AIR2
Temp
°F
WEIGHT IN A
CUBIC FOOT
OF MIXTURE
VOLUME
ft3/lb
WEIGHT OF
THE VAPOR
OF
TOTAL
lb/lb
lb/lb
DRY AIR VAPOR
WEIGHT DRY AIR
OF
OF
lb
lb
lb
DRY AIR MIXTURE
Temp
°F
WEIGHT IN A
CUBIC FOOT
OF MIXTURE
VOLUME
ft3/lb
WEIGHT OF
THE VAPOR
OF
TOTAL
lb/lb
lb/lb
DRY AIR VAPOR
WEIGHT DRY AIR
OF
OF
lb
lb
lb
DRY AIR MIXTURE
81
82
83
84
85
.07081
.07059
.07038
.07015
.06993
.001629
.001680
.001733
.001785
.001840
.07244
.07227
.07211
.07193
.07177
14.12
14.16
14.21
14.26
14.30
.02301
.02380
.02462
.02545
.02631
.02249
.02325
.02403
.02482
.02566
116
117
118
119
120
.06186
.06154
.06124
.06092
.06060
.004427
.004548
.004669
.004794
.004921
.06629
.06609
.06591
.06571
.06552
16.16
16.24
16.32
16.41
16.50
.07157
.07390
.07625
.07869
.08121
.06678
.06882
.07084
.07296
.07511
86
87
88
89
90
.06970
.06947
.06925
.06902
.06880
.001898
.001954
.002014
.002072
.002139
.07160
.07142
.07126
.07109
.07094
14.34
14.39
14.44
14.48
14.53
.02723
.02813
.02908
.03002
.03109
.02651
.02736
.02826
.02915
.03015
121
122
123
124
125
.06027
.05995
.05960
.05927
.05892
.005049
.005183
.005319
.005456
.005598
.06532
.06513
.06492
.06473
.06452
16.58
16.68
16.77
16.87
16.96
.08376
.08646
.08925
.09204
.09502
.07729
.07958
.08194
.08428
.08677
91
92
93
94
95
.06855
.06832
.06809
.06785
.06760
.002201
.002267
.002334
.002404
.002474
.07075
.07058
.07042
.07025
.07007
14.58
14.63
14.69
14.73
14.79
.03211
.03318
.03428
.03543
.03660
.03111
.03212
.03314
.03422
.03531
130
135
140
145
150
.05713
.05524
.05319
.05100
.04865
.006355
.007195
.008128
.009162
.010303
.06349
.06244
.06132
.06016
.05895
17.49
18.10
18.79
19.60
20.55
.11125
.13026
.15280
.17966
.21178
.10010
.11523
.13255
.15230
.17478
96
97
98
99
100
.06736
.06711
.06688
.06660
.06634
.002546
.002620
.002692
.002770
.002853
.06991
.06973
.06957
.06931
.06919
14.84
14.90
14.95
15.01
15.07
.03780
.03904
.04025
.04159
.04300
.03642
.03757
.03870
.03993
.04124
155
160
165
170
175
.04612
.04340
.04048
.03734
.03398
.011547
.012937
.014436
.016118
.017926
.05767
.05634
.05492
.05346
.05191
21.67
23.03
24.69
26.77
29.43
.25038
.29810
.35660
.43168
.52750
.20022
.22962
.26285
.30150
.34530
101
102
103
104
105
.06610
.06583
.06557
.06530
.06504
.002937
.003019
.003106
.003193
.003283
.06904
.06885
.06868
.06849
.06832
15.12
15.18
15.25
15.31
15.37
.04443
.04586
.04737
.04890
.05048
.04255
.04385
.04523
.04662
.04806
180
185
190
195
200
.03035
.02645
.02228
.01779
.01297
.019905
.022062
.024393
.026957
.029730
.05036
.04851
.04667
.04475
.04270
32.94
37.78
44.85
56.20
77.11
.65580
.83410
1.0948
1.5153
2.2923
.39525
.45425
.52270
.60240
.69660
106
107
108
109
110
.06477
.06451
.06421
.06394
.06364
.003375
.003470
.003568
.003666
.003766
.06814
.06798
.06778
.06761
.06741
15.44
15.50
15.57
15.64
15.71
.05212
.05379
.05556
.05734
.05917
.04953
.05105
.05264
.05422
.05587
205
210
212
.00782
.00232
.00000
.032715
.035942
.037298
.04064
.03836
.03730
127.9
431.0
____
4.1838
15.493
Inf.
.80500
.93700
1.0000
111
112
113
114
115
.06336
.06306
.06278
.06247
.06216
.003872
.003978
.004085
.004199
.004311
.06723
.06704
.06686
.06667
.06647
15.78
15.85
15.93
16.00
16.08
.06111
.06308
.06507
.06722
.06935
.05760
.05934
.06110
.06299
.06486
Figure N.4 - Weights of Air, Water Vapor, and Saturated Mixture of Air and
Water Vapor at Different Temperatures and 29.92 in. Hg
2. Jorgensen, op. cit., pp 15-17
Reprinted by Permission
Field Performance Measurement | 265
Density of Saturated Air for Various Barometric Conditions - lbm/ft3
Dry-Bulb
Temp. °F
Barometric Pressure in. Hg
Approximate
average
Increase in increase in
density per density per
°F wet-bulb
0.1 in.
pressure depression
28.5
29.0
29.5
30.0
30.5
31.0
30
31
32
33
34
.07703
.07687
.07671
.07654
.07638
.07839
.07822
.07806
.07789
.07772
.07974
.07957
.07940
.07924
.07907
.08110
.08093
.08075
.08058
.08041
.08245
.08228
.08210
.08193
.08175
.08380
.08363
.08345
.08327
.08310
.00027
.00027
.00027
.00027
.00027
.000017
.000017
.000017
.000018
.000018
35
36
37
38
39
.07621
.07605
.07589
.07573
.07557
.07756
.07739
.07723
.07706
.07690
.07890
.07873
.07856
.07840
.07823
.08024
.07807
.07990
.07973
.07956
.08158
.08141
.08123
.08106
.08089
.08292
.08274
.08257
.08239
.08222
.00027
.00027
.00027
.00027
.00027
.000018
.000018
.000019
.000019
.000019
40
41
42
43
44
.07541
.07525
.07509
.07493
.07477
.07674
.07657
.07641
.07625
.07609
.07806
.07790
.07773
.07757
.07740
.07939
.07922
.09705
.07889
.07872
.08072
.08055
.08038
.08021
.08004
.08205
.08187
.08170
.08153
.08135
.00027
.00026
.00026
.00026
.00026
.000019
.000020
.000020
.000020
.000020
45
46
47
48
49
.07461
.07445
.07429
.07413
.07397
.07592
.07576
.07560
.07544
.07528
.07724
.07707
.07691
.07674
.07658
.07855
.07838
.07822
.07805
.07788
.07986
.07970
.07953
.07936
.07919
.08118
.08101
.08084
.08066
.08049
.00026
.00026
.00026
.00026
.00026
.000020
.000021
.000021
.000021
.000022
50
51
52
53
54
.07381
.07366
.07350
.07334
.07318
.07512
.07496
.07479
.07464
.07447
.07642
.07625
.07609
.07593
.07576
.07772
.07755
.07739
.07722
.07706
.07902
.07885
.07868
.07852
.07835
.08032
.08015
.07998
.07981
.07964
.00026
.00026
.00026
.00026
.00026
.000022
.000022
.000023
.000023
.000023
55
56
57
58
59
.07302
.07287
.07271
.07255
.07240
.07431
.07415
.07399
.07383
.07367
.07560
.07544
.07528
.07512
.07495
.07689
.07673
.07656
.07640
.07623
.07818
.07801
.07784
.07768
.07751
.07947
.07930
.07913
.07896
.07879
.00026
.00026
.00026
.00026
.00026
.000024
.000024
.000025
.000025
.000025
60
61
62
63
64
.07224
.07208
.07193
.07177
.07161
.07352
.07336
.07320
.07304
.07288
.07479
.07463
.07447
.07430
.07414
.07607
.07590
.07574
.07557
.07541
.07734
.07718
.07701
.07684
.07668
.07862
.07845
.07828
.07811
.07794
.00026
.00026
.00026
.00026
.00026
.000026
.000026
.000027
.000027
.000028
Note: Approximate average decrease in density per 0.1°F rise in dry-bulb temperature equals .000017 lbm/ft3.
Figure N.5 - Psychrometric Density Table (I-P)
266 | Field Performance Measurement
Psychrometric Density Table (I-P)
Density of Saturated Air for Various Barometric Conditions - lbm/ft3
Dry-Bulb
Temp. °F
Barometric Pressure in. Hg
Approximate
average
Increase in increase in
density per density per
°F wet-bulb
0.1 in.
pressure depression
28.5
29.0
29.5
30.0
30.5
31.0
65
66
67
68
69
.07145
.07130
.07114
.07098
.07083
.07272
.07256
.07240
.07224
.07208
.07398
.07382
.07366
.07350
.07333
.07525
.07508
.07492
.07475
.07459
.07651
.07634
.07618
.07601
.07584
.07770
.07760
.07744
.07727
.07710
.00026
.00026
.00026
.00026
.00026
.000028
.000029
.000029
.000030
.000030
70
71
72
73
74
.07067
.07051
.07035
.07020
.07004
.07192
.07176
.07160
.07144
.07128
.07317
.07301
.07285
.07268
.07252
.07442
.07426
.07410
.07393
.07377
.07568
.07551
.07534
.07517
.07501
.07693
.07676
.07659
.07642
.07625
.00026
.00025
.00025
.00025
.00025
.000031
.000031
.000032
.000033
.000033
75
76
77
78
79
.06988
.06972
.06956
.06940
.06925
.07112
.07096
.07080
.07064
.07048
.07236
.07220
.07203
.07187
.07171
.07360
.07343
.07327
.07310
.07294
.07484
.07467
.07451
.07434
.07417
.07603
.07591
.07574
.07557
.07540
.00025
.00025
.00025
.00025
.00025
.000034
.000034
.000035
.000036
.000036
80
81
82
83
84
.06909
.06893
.06877
.06861
.06845
.07032
.07015
.07000
.06983
.06967
.07155
.07138
.07122
.07105
.07089
.07277
.07261
.07244
.07227
.07211
.07400
.07383
.07366
.07349
.07333
.07523
.07506
.07489
.07472
.07454
.00025
.00025
.00024
.00024
.00024
.000037
.000038
.000039
.000039
.000040
85
86
87
88
89
.06829
.06812
.06796
.06780
.06764
.06950
.06934
.06917
.06901
.06885
.07072
.07056
.07039
.07022
.07005
.07194
.07177
.07160
.07143
.07126
.07316
.07299
.07281
.07264
.07247
.07437
.07420
.07403
.07385
.07368
.00024
.00024
.00024
.00024
.00024
.000041
.000042
.000043
.000043
.000044
90
91
92
93
94
.06748
.06731
.06715
.06698
.06682
.06868
.06852
.06835
.06818
.06801
.06989
.06972
.06955
.06938
.06921
.07109
.07092
.07075
.07058
.07041
.07230
.07213
.07195
.07178
.07161
.07351
.07333
.07316
.07298
.07280
.00024
.00024
.00024
.00024
.00024
.000045
.000046
.000047
.000048
.000049
95
96
97
98
99
.06665
.06648
.06632
.06615
.06598
.06785
.06768
.06751
.06734
.06717
.06904
.06887
.06870
.06853
.06835
.07024
.07006
.06989
.06972
.06954
.07143
.07126
.07108
.01091
.07073
.07263
.07245
.07227
.07209
.07191
.00024
.00024
.00024
.00024
.00024
.000050
.000051
.000052
.000053
.000054
100
.06581
.06700
.06818
.06937
.07055
.07174
.00024
.000055
Note: Approximate average decrease in density per 0.1°F rise in dry-bulb temperature equals .000017 lbm/ft3.
Figure N.5 - Psychrometric Density Table (I-P)
Field Performance Measurement | 267
FUEL
FLUE GAS DENSITY
lbm/ft3
COAL
0.078
OIL
0.075
NATURAL GAS
0.0725
BAGASSE
0.070
BLAST FURNACE GAS
0.076
LIGNITE
0.073
WOOD
0.070
The above densities at 70°F and 29.92 in. Hg are based on average fuel analyses and moisture contents
Figure N.6 - Typical Densities for Various Flue Gases
268 | Field Performance Measurement
Annex P. Diffusion at Fan Outlets
BLAST AREA
DISCHARGE DUCT
CUTOFF
OUTLET AREA
25%
50%
75%
CENTRIFUGAL FAN
100% EFFECTIVE DUCT LENGTH
AXIAL FAN
To calculate 100% effective duct length, assume a minimum of 2½ duct diameters for 2500 fpm or less. Add 1 duct
diameter for each additional 1000 fpm.
Example: 5000 fpm = 5 equivalent duct diameters
If the duct is rectangular, with side dimensions equal to a and b, the equivalent duct diameter is equal to (4ab/π)0.5
Figure P.1 - Controlled Diffusion and Establishment of a Uniform
Velocity Profile in a Straight Length of Outlet Duct
Field Performance Measurement | 269
Annex R. Terminology for Fans and Air Handling Units
CASING
BACKPLATE
RIM
INLET
HUB
MOTOR
GUIDE VANE
BLADE
IMPELLER
INLET BELL
Tubular Centrifugal Fan - Direct Drive
CASING
BLADE
DIFFUSER
HUB
MOTOR
IMPELLER
CASING
Tubeaxial Fan-Direct Drive
(Impeller Downstream)
BEARING CASING
BELT TUBE
BLADE
HUB
GUIDE VANE
Vaneaxial Fan-Belt Drive
IMPELLER
INLET BOX
BEARINGS
FAN
CASING
GUIDE VANES
MECHANISM FOR
CONTROLLING
BLADE ANGLE
INNER CYLINDER
IMPELLER
DIFFUSER
Vaneaxial Mechanical Draft Fan
Figure R.1 - Common Terminology for Axial and Tubular Centrifugal Fans
270 | Field Performance Measurement
HOUSING
DIVERTER
CU
TO
FF
CENTER PLATE
BLAST AREA
DISCHARGE
OUTLET AREA
SIDE SHEET
BACKPLATE
FF
BLADE
TO
CU
INLET
SCROLL
IMPELLER
FRAME
RIM
BEARING
SUPPORT
INLET COLLAR
Figure R.2 - Common Terminology for Centrifugal Fan
Field Performance Measurement | 271
Figure R.3 - Common Terminology for Centrifugal Fan Appurtenances
272 | Field Performance Measurement
HEATING AND VENTILATING DRAW-THROUGH UNIT
FS
BELT GUARD
FS CS
EXT
F & BP
MB
FB
INT
F & BP
HC
MB
FB
AS
+
+
HEATING AND VENTILATING BLOW-THROUGH UNIT
ZONE DAMPERS
FS
HC
BYPASS
COLD DECK
+
+
HOT DECK
+
+
AIR-CONDITIONING DRAW-THROUGH UNIT
FS
AS
MB
FB
CC
SS
HC
ELIM
+
+
+
MB
FB
+
+
+
+
+
DRIP TRAY
AIR-CONDITIONING BLOW-THROUGH UNIT
DIFFUSER
PLATE
ZONE DAMPERS
FS
CC
+
+
+
+
EXTERNAL FACE AND BYPASS DAMPER
INTERNAL FACE AND BYPASS DAMPER
ELIMINATORS
FS
FB
MB
SS
+
HOT DECK
+
HC
HC
FB
MB
+
COLD DECK
CC
+
FLEXIBLE CONNECTION
AS
CS
CC
HC
ACCESS SECTION
COIL SECTION
COOLING COIL
HEATING COIL
EXT F & BP
INT F & BP
ELIM
FAN SECTION
FILTER BOX
MIXING BOX
SPRAY SECTION
Figure R.4 - Common Terminology for Central Station Air-Handling Units
Field Performance Measurement | 273
Annex S. Typical Format for Field Test Data Sheet
FIELD TEST DATA SHEET
JOB DESCRIPTION: Location, User, Contractor, Engineer, . . . . .
FAN DESCRIPTION: Mfgr., Size, Type, Ident. No., . . . . .
MOTOR DESCRIPTION: Mfgr., Nameplate Data (Ident. No., hp, volts, FLA, . . . ), Performance Data
Reference, . . . . .
DRIVE DESCRIPTION: Type, Mfgr., Ident. No., Size, . . . . .
REFERENCE DRAWINGS OR
SKETCHES OF INSTALLATION: System Configuration with Dimensions, Measurement Plane
Locations, . . . . .
MEASUREMENTS
AMBIENT DATA: Barometric Pressure, Dry-Bulb Temp., Wet-Bulb Temp, . . . . .
MOTOR DATA: volts, amps, watts, rpm, . . . . .
FAN SPEED
GAS DENSITY DATA:
GAS TEMPERATURES AT MEASUREMENT PLANES:
READING
Ps1 or Ps4
Ps2 or Ps5
Ps3
Pv3
Pv3
1
2
3
4
5
•
•
•
•
n
TOTAL
AVERAGE
CALCULATIONS: (Refer to the various sections of this publication for the appropriate calculation
procedures.)
Figure S.1 - Typical Format for Field Test Data Sheet
274 | Field Performance Measurement
Annex T. Uncertainty Analysis
T.1 Introduction
In an attempt to determine the range of uncertainties
likely to be encountered in field testing of fans, a
statistical uncertainty analysis was undertaken.
Maximum and minimum uncertainties were assigned
to each quantity to be measured based on the degree
of difficulty in measuring the quantity, the previously
specified accuracies of instruments and the
conditions expected to be encountered in field
testing. These individual maximum and minimum
uncertainties were then combined statistically to
arrive at the probable range of overall uncertainties
for the fan flow rate, fan static pressure, and fan
power input. It would be unlikely, however, that any
particular field installation would have all minimum or
all maximum uncertainties occurring simultaneously.
Therefore, an agreement by the parties as to
acceptable measurement tolerances for a given
installation should be established prior to testing.
In Type A tests, it may be sufficient to accept the
results of any field test without consideration of the
probable uncertainties in the results. For Type B and
Type C tests, it may be necessary to calculate the
uncertainties. To do this, each measured quantity is
assigned an estimated uncertainty by agreement of
the parties involved and the overall uncertainty is
calculated as outlined in this annex.
T.2 General
This analysis is based on the assumption that fan
perfomance can be treated as a statistical quantity
and that the performances derived from repeated
tests would have a normal distribution. The most
probable performance would, therefore, be the mean
results based on repeated observations at each point
of operation. Only one set of observations is
specified in this publication. This analysis deals,
therefore, with the probable uncertainty in the results
obtained from a single set of observations.
The results of a fan field performance test for a single
point of operation are a combination of variables
which are normally presented graphically. Test results
will be considered to be the fan static pressure
versus flow rate and fan power input versus flow rate.
The uncertainty in results will be expressed in terms
of fan flow rate, fan static pressure, and fan power
input.
The accuracies specified in this publication are based
upon two standard deviations. This means that there
should be a 95% probability that the actual
uncertainties will be less than the specified value.
This applies only to random uncertainties. Systematic
uncertainties should be eliminated by the use of
properly calibrated test instruments. This analysis
considers only the uncertainties inherent in testing.
This publication specifies uncertainties in percent.
These are, of course, per unit uncertainties,
multiplied by 100. Absolute uncertainties which bear
the units of the quantity being measured or
calculated, are equal to the per unit uncertainty
multiplied by the measured or calculated quantity.
Since the tolerance on measured values is specified
on the basis of 95% confidence limits, the actual
deviations in results will be less than the calculated
deviations 95% of the time.
For the purposes of a field test, an uncertainty range
will be defined with minimum and maximum values.
This range of possible uncertainty is necessary to
cover the varying degrees of difficulty encountered in
performing tests in field installations. Field test
conditions range from near ideal to near impossible.
T.3 Symbols
In the analysis that follows, certain symbols and
notations are used in addition to those shown in
Annex Q.
Symbol
Quantity
ex
∆X
R
Per Unit Uncertainty in X
Absolute Uncertainty in X
Gas Constant (ft-lb/lbm —°R)
Subscript
Description
A
b
d
f
g
h
H
N
P
Q
w
x
ρ
area
Barometric Pressure
Dry-bulb Temperature
Velocity Pressure
Static Pressure
Power Input
Fan Power Input
Fan Speed
Fan Static Pressure
Fan Flow Rate
Wet-bulb Depression
Generalized Quantity (A, b, ..., ρ)
Density
T.4 Measurement uncertainties
The various measurement uncertainty ranges used in
this publication are listed below. The considerations
that led to their adoption include difficulties in field
testing generally not encountered in laboratory
testing.
Field Performance Measurement | 275
T.4.1 Barometric pressure. The estimated
uncertainty in measuring barometric pressure is
between 0.3% minimum and 0.7% maximum.
eb = 0.003 (min) to 0.007 (max)
Barometric pressure is generally obtained by
portable aneroid barometer, on-site barometer
(mercury or aneroid) or by use of data obtained from
a nearby airport. The uncertainty range above is
estimated based on the use of portable or on-site
instrumentation and applicable corrections.
T.4.2 Dry-bulb temperature. The estimated
uncertainty in measuring dry-bulb temperature is
between 0.5% of absolute temperature minimum and
2.0% of absolute temperature maximum.
ed = 0.005 (min) to 0.02 (max)
The estimated uncertainty range is based on a broad
temeprature range and the likelihood of stratification.
T.4.3 Web-bulb depression. The estimated
uncertainty in measuring wet-bulb depression is
between 5°F minimum and 10°F maximum.
ew = 5/(td - tw) (min) to 10/(td - tw) (max)
The estimated uncertainty range is based on a broad
temperature range with the associated difficulties in
determining wet-bulb readings at high or low
temperatures and the likelihood of stratification.
T.4.4 Fan speed. The estimated uncertainty in
measuring fan speed is between 0.5% minimum and
1.0% maximum.
eN = 0.005 (min) to 0.01 (max)
The uncertainty range in fan speed is estimated on
the basis of portable instrumentation accuracy and
an allowance for fluctuation in fan speed.
T.4.5 Power input. The estimated uncertainty in
measuring power input is betwen 3.0% minimum and
7.0% maximum.
eh = 0.03 (min) to 0.07 (max)
The estimated uncertainty range is based on the
various measurement methods and their respective
accuracies, estimated drive losses, and the broad
horsepower range encountered in the field.
T.4.6 Pitot traverse. A properly performed field
traverse is estimated to have an accuracy of 1.5%
minimum to 7.5% maximum.
276 | Field Performance Measurement
ec = 0.015 (min) to 0.075 (max)
The uncertainty range in the Pitot traverse is
estimated on the basis of traverse location, broad
range of duct sizes, nonuniform velocity profiles, and
turbulence.
T.4.7 Flow measurement area. The estimated
uncertainty in the flow measurement area is between
1.0% minimum to 2.0% maximum.
eA = 0.010 (min) to 0.020 (max)
The estimated uncertainty is based on a broad range
of duct sizes, accessibility, and the rigidity of ducts
under pressure.
T.4.8 Velocity pressure. An allowance of 2.0%
minimum to 5.0% maximum of the reading is
estimated for the mental averaging performed on a
fluctuating reading. An allowance of 1.0% minimum
to 2.0% maximum of the reading is estimated for
calibrated manometer uncertainty and relocation of
the instrument after calibration. In addition, an
allowance of 0.5% minimum to 10.0% maximum of
the reading is estimated for instrument precision. No
allowance is included for yaw on the assumption that
the Pitot-static tube is aligned within 10 degrees of
streamlines. A combined uncertainty can be written
as:
ef (min) = [(0.02)2 + (0.01)2 + (0.005)2]0.5
= 0.0229
ef (max) = [(0.05)2 + (0.02)2 + (0.10)2]0.5
= 0.1136
T.4.9 Static pressure. An allowance of 1.0%
minimum to 5.0% maximum of the reading is
estimated for the mental averaging performed on a
fluctuating reading. An allowance of 1.0% minimum
to 2.0% maximum of the reading is estimated for
calibrated manometer uncertainty and relocation of
the instrument after . In addition, a tolerance of 10%
minimum to 20.0% maximum of the fan velocity
pressure should cover the influence of Pitot-static
tube yaw or velocity influence on static pressure taps
and other possible effects. A combined uncertainty
can be written as:
eg (min) = {(0.01)2 + (0.01)2 + (0.005)2 +
[0.1 Pv/(Ps2 - Ps1)]2}0.5
= {0.000225 + [0.1 Pv/(Ps2 - Ps1)]2}0.5
eg (max) = {(0.05)2 + (0.02)2 + (0.02)2 +
[0.2 Pv/(Ps2 - Ps1)]2}0.5
= {0.0033 + [0.2 Pv/(Ps2 - Ps1)]2}0.5
Where the denominator in the final term in each
equation will involve Ps2 or Ps5 and Ps1 or Ps4,
whichever are measured.
Assuming ∆70.73 and ∆R are both zero:
The estimated uncertainty range is based on an
allowance for fluctuation in the fan-system operation,
lack of ideal measurement locations, turbulence, and
the relocation of instrumentation after calibration.
It can be shown that:
ev2 = [(0.00000725 tw - 0.0000542) ∆(td - tw)]2
Where:
T.5 Combined uncertainties
The uncertainties in the test performance are the
result of using various values, each of which contains
a probable uncertainty. The combined uncertainty for
each of the fan performance variables is given below.
T.5.1 Density. Air density involves the various
psychrometric measurements and the approximate
formula:
ρ=
eρ = (eb2 + ev2 + ed2)0.5
70.73 pbV
R ( t d + 460 )
∆(td - tw) = Absolute uncertainty in wet-bulb depression.
Other methods for determining density are assumed
to have equal accuracy.
T.5.2 Fan flow rate. Fan flow rate directly involves
the area at the flow measuring station, the Pitot
traverse, the square root of the pressure
measurement for flow, and the square root of the
density. Uncertainties in fan speed will produce a
first-power uncertainty in flow rate when making the
fan law conversions. Combining:
eQ = [ec2 + eA2 (ef/2)2 + (eρ/2)2 + eN2]0.5
Where:
V = 1.0 - 0.378 {(pe/pb) - [(td - tw)/2700]}
For random and independant uncertainties in
products, the combined uncertainty is determined as
follows:
∆ρ/ρ = {(∆70.73/70.73)2 + (∆pb/pb)2 + (∆V/V)2 +
(∆R/R)2 + [∆td/(td + 460)]2}0.5
T.5.3 Fan static pressure. Fan static pressure
directly involves static pressure measurements.
Uncertainties in density will produce a first-power
uncertainty in fan static pressure while uncertainties
in fan speed will produce a second-power uncertainty
in fan static pressure when making fan law
conversions. Combining:
ep = [eg2 + eρ2 + (2eN)2]0.5
Table T.1
Measurement
eb
ed**
eW
eN
eh
ec
eA
ef
eg
Minimum
Maximum
0.003
0.005
5/(td - tw)
0.005
0.030
0.015
0.010
0.0229
{0.000225 + [0.1 Pv/(Ps2 - Ps1)]2}0.5
0.007
0.020
10/(td - tw)
0.010
0.070
0.075
0.020
0.1136
{0.0033 + [0.2 Pv/(Ps2 - Ps1)]2}0.5
* These uncertainties do not account for the effect of swirl at the fan inlet. This situation must be corrected in order
to produce acceptable fan-system performance (see Section 5).
** Based on absolute temperature
Field Performance Measurement | 277
In order to simplify the application of this uncertainty
analysis to the results of field tests, the above
equation was developed on the basis of tests in
which static pressure measurements are made at a
single plane, as would be the case in which a fan is
ducted on one side only. However, the equation is
reasonably accurate for all other fan-system
configurations.
Although in most cases the determination of fan static
pressure involves Pv1, the uncertainty in determining
Pv1 is not included in the above equation on the basis
that it normally has a very small effect on the overall
uncertainty in fan static pressure.
For purposes of this publication, eP is applied directly
to Psc, which may include System Effect Factors.
T.5.4 Fan power input. Fan power input directly
involves the power measurement; in addition, when
making fan law conversions, density has a first-power
effect and speed has a third-power effect on fan
power input. Combining:
eH = [eh2 + eρ2 + (3eN)2]0.5
T.6 Summary
The minimum and maximum measurement
uncertainties (See Table T.1) were defined earlier in
Section T.4. Summarizing, the per unit uncertainties
are as shown in Table T.1.
The uncertainty calculations lead to absolute
uncertainties in fan flow rate, fan static pressure, and
fan power input that can be applied directly to the
corresponding test results. The uncertainty results
can then be plotted as rectangles around the test
point. Intersection of the rectangles with the quoted
fan performance within the limitations of a field test.
See the examples in Section T.7.
T.7 Examples
Two examples of the calculation of uncertainties and
the method of comparison with the quoted fan curve
are included in this section. Uncertainty calculations
and comparisons have been developed for Examples
2B and 2C of Annex A. Uncertainty calculations for
Example 2B utilize all minimum uncertainty
tolerances. Uncertainty calculations for Example 2C
utilize all maximum uncertainty tolerances. It would
be unlikely that any field installation would lend itself
to all minimum or all maximum measurement
tolerances. Agreement of the parties as to acceptable
measurement tolerances for a given installation
should be established prior to testing.
278 | Field Performance Measurement
EXAMPLE 1: CALCULATION OF UNCERTAINTIES
IN TEST RESULTS BASED ON MINIMUM
MEASUREMENT UNCERTAINTY
TEST VALUES
Reference: Example 2B in Annex A
SITE MEASUREMENTS
td2 = 91.3°F
tw2 = 70.4°F
Ps1 = -11.4 in. wg
Ps2 = 0.1 in. wg
Pv3 = 1.24 in. wg
A2 = 1.40 ft2
A3 = 1.57 ft2
ρ2 = 0.0714 lbm/ft3
ρ3 = 0.0705 lbm/ft3
CONVERTED RESULTS
Qc = 7114 cfm
Psc = 11.42 in. wg
Hc = 18.90 hp
MEASUREMENT UNCERTAINTIES
Reference: Minimum values per Section T.6
eb = 0.003
ed = 0.005
ew = 5/(td2 - tw2)
eN = 0.005
eh = 0.030
ec = 0.015
eA = 0.010
ef = 0.0229
eg = {0.000225 + [0.1 Pv/(Ps2 - Ps1)]2}0.5
CALCULATIONS
Pv = Pv2
= Pv3 (A3/A2)2 (ρ3/ρ2)
= 1.24 (1.57/1.40)2 (0.0705/0.0714)
= 1.54 in. wg
eg = {0.000225 + [0.1 Pv/(Ps2 - Ps1)]2}0.5
= {0.000225 + [(0.1 × 1.54)/(0.1 + 11.4)]2}0.5
= 0.02011
ev2 = [(0.00000725 tw - 0.0000542) ∆(td - tw)]2
= [(0.00000725 × 70.4 - 0.0000542) 5]2
= 0.00000520
eρ = [eb2 + ev2 + ed2)0.5
= (0.0032 + 0.00000520 + 0.0052)0.5
= 0.006261
eP = [eg2 + eρ2 + (2eN)2]0.5
= [0.020112 + 0.0062612 + (2 × 0.005)2]0.5
= 0.0233
eQ = [ec2 + eA2 + (ef/2)2 + (eρ/2)2 + eN2]0.5
= [0.0152 + 0.0102 + (0.0229/2)2 +
(0.006261/2)2 + 0.0052]0.5
= 0.0222
eH = [eh2 + eρ2 + (3eN)2]0.5
= [0.0302 + 0.0062612 + (3 × 0.005)2]0.5
= 0.0341
∆P = ePPsc
= 0.0233 × 11.42
= 0.27 in. wg
Psc + ∆P = 11.42 + 0.27
= 11.69 in. wg
Psc - ∆P = 11.42 - 0.27
= 11.15 in. wg
∆Q = eQQc
= 0.0222 × 7114
= 158 cfm
Qc + ∆Q = 7114 + 158
= 7272 cfm
Qc - ∆Q = 7114 - 158
= 6956 cfm
∆H = eHHc
= 0.0341 × 18.90
= 0.64 hp
Hc + ∆H = 18.90 + 0.64
= 19.54 hp
Hc - ∆H = 18.90 - 0.64
= 18.26 hp
Field Performance Measurement | 279
GRAPHICAL PRESENTATION
Psc
Psc + ∆P
TEST POINT
MINIMUM UNCERTAINTY RANGE
Ps, FAN STATIC PRESSURE
Psc - ∆P
Qc = 7114 cfm
∆Q = 158 cfm
Psc = 11.42 in. wg
∆P = 0.27 in. wg
Hc = 18.90 hp
∆H = 0.64 hp
Qc + ∆Q
Qc - ∆Q
Qc
QUOTED FAN
PERFORMANCE
CURVES
Q, FAN FLOW RATE
H, FAN POWER INPUT
Hc + ∆H
Hc
Hc - ∆H
Qc + ∆Q
Qc - ∆Q
Qc
Q, FAN FLOW RATE
Figure T.1
280 | Field Performance Measurement
EXAMPLE 2: CALCULATION OF UNCERTAINTIES
IN TEST RESULTS BASED ON MAXIMUM
MEASUREMENT UNCERTAINTIES
TEST VALUES
Reference: Example 2C in Annex A
SITE MEASUREMENTS
eQ = [ec2 + eA2 + (ef/2)2 + (eρ/2)2 + eN2]0.5
= [0.0752 + 0.0202 + (0.1136/2)2
+ (0.02176/2)2 + 0.0102]0.5
= 0.0973
eH = [eh2 + eρ2 + (3eN)2]0.5
= [0.0702 + 0.021762 + (3 × 0.010)2]0.5
= 0.0792
td3 = 86.5°F
tw3 = 75.5°F
Ps4 = -1.57 in. wg
Ps5 = 1.22 in. wg
Pv2 = 0.61 in. wg
∆P = eP Psc
= 0.0780 × 2.54
= 0.20 in. wg
CONVERTED RESULTS
Psc - ∆P = 2.54 - 0.20
= 2.34 in. wg
Psc + ∆P = 2.54 + 0.20
= 2.74 in. wg
Qc = 25964 cfm
Psc = 2.54 in. wg
Hc = 17.11 hp
MEASUREMENT UNCERTAINTIES
Reference: Maximum values per Section T.6
eb = 0.007
ed = 0.020
eW = 10/(td3 - tw3)
eN = 0.010
eh = 0.070
ec = 0.075
eA = 0.020
ef = 0.1136
eg = {0.0033 + [0.2 Pv/(Ps5 - Ps4)]2}0.5
CALCULATIONS
∆Q = eQQc
= 0.0973 × 25964
= 2526 cfm
Qc + ∆Q = 25964 + 2526
= 28490 cfm
Qc - ∆Q = 25964 - 2526
= 23438 cfm
∆H = eHHc
= 0.0792 × 17.11
= 1.36 hp
Hc + ∆H = 17.11 + 1.36
= 18.47 hp
Hc - ∆H = 17.11 - 1.36
= 15.75 hp
eg = {0.0033 + [0.2 Pv/(Ps5 - Ps4)]2}0.5
= {0.0033 + [(0.2 × 0.61)/(1.22 + 1.57)]2}0.5
= 0.07219
ev2 = [(0.00000725 tw - 0.0000542) ∆(td - tw)]2
= [(0.00000725 × 75.5 - 0.0000542) 10]2
= 0.0000243
eρ = (eb2 + ev2 + ed2)0.5
= (0.0072 + 0.0000243 + 0.0202)0.5
= 0.02176
eP = [eg2 + eρ2 + (2eN)2]0.5
= [0.072192 + 0.021762 + (2 × 0.010)2]0.5
= 0.0780
Field Performance Measurement | 281
GRAPHICAL PRESENTATION
TEST POINT
MAXIMUM UNCERTAINTY RANGE
Qc = 25964 cfm
∆Q = 2526 cfm
Psc + ∆P
Psc = 2.54 in. wg
∆P = 0.20 in. wg
Ps, FAN STATIC PRESSURE
Psc
Hc = 17.11 hp
∆H = 1.36 hp
Psc - ∆P
Qc - ∆Q
Qc + ∆Q
Qsc
H, FAN POWER INPUT
Q, FAN FLOW RATE
QUOTED FAN
PERFORMANCE
CURVES
Hc + ∆H
Hc
Hc - ∆H
Qc - ∆Q
Qc + ∆Q
Qsc
Q, FAN FLOW RATE
Figure T.2
282 | Field Performance Measurement
AIR MOVEMENT AND CONTROL
ASSOCIATION INTERNATIONAL, INC.
30 West University Drive
Arlington Heights, IL 60004-1893 U.S.A.
Tel: (847) 394-0150
E-Mail : info@amca.org
Fax: (847) 253-0088
Web: www.amca.org
The Air Movement and control Association International, Inc. is a not-for-profit international association of the
world’s manufacturers of related air system equipment primarily, but limited to: fans, louvers, dampers, air
curtains, airflow measurement stations, acoustic attenuators, and other air system components for the industrial,
commercial and residential markets.