AMCA Publication 201 02 R2007

The International Authority on Air System Components

AIR MOVEMENT AND CONTROLASSOCIATION INTERNATIONAL, INC.

AMCAPublication 201-02

Fans and Systems

(R2007)

AMCA PUBLICATION 201-02 (R2007)

Fans and Systems

Air Movement and Control Association International, Inc.

30 West University Drive

Arlington Heights, IL 60004-1893

© 2007 by Air Movement and Control Association International, Inc.

All rights reserved. Reproduction or translation of any part of this work beyond that permitted by Sections 107 and

108 of the United States Copyright Act without the permission of the copyright owner is unlawful. Requests for

permission or further information should be addressed to the Executive Director, Air Movement and Control

Association International, Inc. at 30 West University Drive, Arlington Heights, IL 60004-1893 U.S.A.

Forward

ANSI/AMCA Standard 210 Laboratory Methods of Testing Fans for Aerodynamic Performance Rating, provides abasis for accurately rating the performance of fans when tested under standardized laboratory conditions. Theactual performance of a fan when installed in an air moving system will sometimes be different from the fanperformance as measured in the laboratory. The difference in performance between the laboratory and the fieldinstallation can sometimes be attributed to the interaction of the fan and the duct system, i.e., duct system designcan diminish the usable output of the fan.

AMCA Publication 201 Fans and Systems, introduced the concept of System Effect Factor to the air movingindustry. The System Effect Factor quantifies the duct system design effect on performance. The System EffectFactor has been widely accepted since its inception in 1973. It must be remembered, however, that the "factors"provided are approximations as it is prohibitive to test all fan types and all duct system configurations. The majorrevision to this edition of AMCA Publication 201 Fans and Systems, is a change to the use of SI units of measure,with Inch-Pound units being given secondary consideration.

AMCA 201 Review Committee

Bill Smiley The Trane Company / LaCrosse

James L. Smith Aerovent, A Twin City Fan Company

Tung Nguyen Emerson Ventilation Products

Patrick Chinoda Hartzell Fan, Inc.

Rick Bursh Illinois Blower, Inc.

Sutton G. Page Austin Air Balancing Corp.

Paul R. Saxon AMCA Staff

Disclaimer

AMCA International uses its best efforts to produce standards for the benefit of the industry and the public in lightof available information and accepted industry practices. However, AMCA International does not guarantee, certifyor assure the safety or performance of any products, components or systems tested, designed, installed oroperated in accordance with AMCA International standards or that any tests conducted under its standards will benon-hazardous or free from risk.

Objections to AMCA Standards and Certifications Programs

Air Movement and Control Association International, Inc. will consider and decide all written complaints regardingits standards, certification programs, or interpretations thereof. For information on procedures for submitting andhandling complaints, write to:

Air Movement and Control Association International30 West University DriveArlington Heights, IL 60004-1893 U.S.A.

or

AMCA International, Incorporatedc/o Federation of Environmental Trade Associations2 Waltham Court, Milley Lane, Hare HatchReading, BerkshireRG10 9TH United Kingdom

Related AMCA Standards and Publications

Publication 200 AIR SYSTEMS

System Pressure Losses

Fan Performance Characteristics

System Effect

System Design Tolerances

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.

Publication 201 FANS AND SYSTEMS

Fan Testing and Rating

The Fan "Laws"

Air Systems

Fan and System Interaction

System Effect Factors

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.

Publication 202 TROUBLESHOOTING

System Checklist

Fan Manufacturer's Analysis

Master Troubleshooting Appendices

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.

Publication 203 FIELD PERFORMANCE MEASUREMENTS OF FAN SYSTEMS

Acceptance Tests

Test Methods and Instruments

Precautions

Limitations and Expected Accuracies

Calculations

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.

TABLE OF CONTENTS

1. Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1

1.1 Purpose . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1

1.2 Some limitations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1

2. Symbols and Subscripts . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1

2.1 Symbols and subscripted symbols . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1

2.2 Subscripts . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1

3. Fan Testing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1

3.1 ANSI/AMCA Standard 210 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1

3.2 Ducted outlet fan tests . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .3

3.3 Free inlet, free outlet fan tests . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4

3.4 Obstructed inlets and outlets . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4

4. Fan Ratings . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4

4.1 The Fan Laws . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4

4.2 Limitations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4

4.3 Fan performance curves . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9

5. Catalog Performance Tables . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .13

5.1 Type A: Free inlet, free outlet fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .13

5.2 Ducted fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .13

6. Air Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .16

6.1 The system . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .16

6.2 Component losses . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .16

6.3 The system curve . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .17

6.4 Interaction of system curve and fan performance curve . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .18

6.5 Effect of changes in speed . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .18

6.6 Effect of density on system resistance . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .19

6.7 Fan and system interaction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .21

6.8 Effects of errors in estimating system resistance . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .21

6.9 Safety factors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .22

6.10 Deficient fan/system performance . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .23

6.11 Precautions to prevent deficient performance . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .23

6.12 System effect . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .23

7. System Effect Factor (SEF) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .24

7.1 System Effect Curves . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .24

7.2 Power determination . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .29

8. Outlet System Effect Factors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .29

8.1 Outlet ducts . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .29

8.2 Outlet diffusers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .30

8.3 Outlet duct elbows . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .31

8.4 Turning vanes . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .35

8.5 Volume control dampers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .35

8.6 Duct branches . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .37

9. Inlet System Effect Factors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .38

9.1 Inlet ducts . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .38

9.2 Inlet duct elbows . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .38

9.3 Inlet vortex (spin or swirl) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .40

9.4 Inlet turning vanes . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .44

9.5 Airflow straighteners . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .44

9.6 Enclosures (plenum and cabinet effects) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .46

9.7 Obstructed inlets . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .47

10. Effects of Factory Supplied Accessories . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .49

10.1 Bearing and supports in fan inlet . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .50

10.2 Drive guards obstructing fan inlet . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .50

10.3 Belt tube in axial fan inlet or outlet . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .50

10.4 Inlet box . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .50

10.5 Inlet box dampers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .50

10.6 Variable inlet vane (VIV) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .51

Annex A. SI / I-P Conversion Table (Informative) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .52

Annex B. Dual Fan Systems - Series and Parallel . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .53

B.1 Fans operating in series . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .53

B.2 Fans operating in parallel . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .53

Annex C. Definitions and Terminology . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .55

C.1 The air . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .55

C.2 The fan . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .55

C.3 The system . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .58

Annex D. Examples of the Convertibility of Energy from Velocity

Pressure to Static Pressure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .62

D.1 Example of fan (tested with free inlet, ducted outlet) applied to a

duct system . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .62

D.2 Example of fan (tested with free inlet, ducted outlet), connected to a

duct system and then a plenum . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .63

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) . . . . . . . . . . . . . . . . . . .65

D.4 Example of fan used to exhaust with obstruction in inlet, inlet elbow,

inlet duct, free outlet . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .66

Annex E. References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .69

AMCA INTERNATIONAL, INC. AMCA 201-02 (R2007)

Fans and Systems

1. Introduction

ANSI/AMCA 210 Laboratory Methods of Testing FansFor Aerodynamic Performance Rating, offers the

system design engineer guidance as to how the fan

was tested and rated. AMCA Publication 201 Fansand 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 Manualincludes 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 Factorsgiven 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.

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.

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 Fansfor Aerodynamic Performance Rating. The purpose

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

1

Table 2.1 - Symbols and Subscripted Symbols

UNITS OF MEASURE

SYMBOL DESCRIPTION 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 Dimensionless

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

Table 2.2 - Subscripts

SUBSCRIPT DESCRIPTION

a Atmospheric conditions

c Converted Value

x Plane 0, 1, 2, ...as appropriate

1 Fan Inlet Plane

2 Fan Outlet Plane

3 Pitot Traverse Plane

5 Plane 5 (nozzle inlet station in chamber)

6 Plane 6 (nozzle discharge station in chamber)

8 Plane 8 (inlet chamber measurement station)

AMCA 201-02 (R2007)

2

TransitionPiece

Straightener

1 2

FOR FAN INSTALLATION TYPES:

B: Free Inlet, Ducted Outlet D: Ducted Inlet, Ducted Outlet

Figure 3.2 - Pitot Traverse in Outlet Duct

AMCA 201-02 (R2007)

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,

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 angle of the transition between the test duct and

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 installationwhere uniform airflow conditions do not exist, thefan'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.

3

4

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

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

AMCA 201-02 (R2007)

5

AMCA 201-02 (R2007)

PL 2PL X

PL 2 PL X

OUTLET AREA

BLAST AREA

CENTRIFUGAL FAN

AXIAL FAN

CUTOFF

DISCHARGE DUCT

PL.5 PL.6 PL.8 PL.1 PL.2

SETTLINGMEANS

VARIABLESUPPLYSYSTEM

SETTLINGMEANS(See note 4)

FAN

0.1 M MIN.

0.5 M MIN.

0.2 M MIN.0.3 M MIN.

P t8PP s5

M

0.2MMIN.

38mm ±6mm(1.5in. ±0.25 in.)

0.5MMIN.

td2

td3

AIRFLOW

Figure 3.3 - Controlled Diffusion and Establishment of a Uniform Velocity

Profile in a Straight Length of Outlet Duct

Figure 3.4 - Inlet Chamber Setup - Multiple Nozzles in Chamber

(ANSI/AMCA 210-99, Figure 15)

AMCA International, Inc. | 30 W. University Dr. | Arlington Heights, IL, 60004-1893 | U.S.A

ANSI/AMCA Standard 99-2404-03 Page 1 of 2

AMCA Drive

Arrangement

ISO 13349

Drive

Arrangement

Description Fan ConfigurationAlternative Fan

Configuration

1 SWSI 1 or

12 (Arr. 1 with

sub-base)

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 2 For belt or direct drive.

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)


Impeller mounted on shaft

between bearings supported by

the fan casing.


on independent pedestals, with


3 DWDI 6 or

18 (Arr. 6 with

sub-base)



between bearings supported by

the fan casing.


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.





Motor flange mounted to

casing.


Drive Arrangements for Centrifugal FansAn American National Standard - Approved by ANSI on April 17, 2003

Figure 3.5 - AMCA Standard 99-2404 / Page 1

AMCA 201-02 (R2007)

6



AMCA Drive

Arrangement

ISO 13349

Drive

Arrangement


Configuration

7 SWSI 7 For coupling drive.

Generally the same as Arr. 3,

with base for the prime mover.


on independent pedestals with


7DWDI 17

(Arr. 6 with

base for motor)

For coupling drive.

Generally the same as Arr. 3











9 SWSI 9 For belt drive.


bearings mounted on pedestal

base.

Motor mounted on the outside

of the bearing base.


10 SWSI 10 For belt drive.


with motor mounted inside of

the bearing pedestal.



AMCA 201-02 AMCA 201-02 (R2007)

7



Drive Arrangements for Axial FansAn American National Standard - Approved by ANSI on June 10, 2003

AMCA Drive

Arrangement

ISO 13349

Drive

Arrangement


Configuration

1 1

12 (Arr. 1 with

sub-base)



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)



between bearings on internal

supports.

Alternative: Fan can be

supplied with inlet box, and/or

discharge evasé.

4 4 For direct drive.



Motor mounted on base or

integrally mounted.

Alternative: With inlet box

and/or with discharge evasé.

M MM M





and/or discharge evasé.

M M




Alternative: Single stage or two

stage fans can be supplied with

inlet box and/or discharge

evasé.

M M

9 9 For belt drive.

Generally same as Arr. 1 with

motor mounted on fan casing,

and/or an integral base.


and/or discharge evasé

M

Note: All fan orientations may be horizontal or vertical


AMCA 201-02 (R2007)

8

AMCA 201-02 (R2007)

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 × 10-3 Pa•s (5°C to 38°C) (SI)

(1.22 × 10-05 lbm/ft•s (40°F to 100°F)) (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

calculated using the proper specific heat ratio for the

gases being handled.

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 Ps

Fan Total Pressure Pt

Fan Power HFan Static Efficiency ηs

Fan Total Efficiency 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.

9

SIZE 30 FAN AT N RPM

OPERATION ATSTANDARD DENSITY

PR

ES

SU

RE

, P

PO

WE

R, H

0

10

20

30

40

50

60

70

80

90

100

AIRFLOW, Q

Pt

Ps

η t

η s

H EF

FIC

IEN

CY, η

PE

RC

EN

T

Figure 4.1 - Fan Performance Curve at N RPM

AMCA 201-02 (R2007)

10

AMCA 201-02 (R2007)

11

TYPE IMPELLER DESIGN HOUSING DESIGN

AIR

FOIL

BA

CK

WA

RD

-IN

CLI

NE

DB

AC

KW

AR

D-

CU

RV

ED

RA

DIA

LFO

RW

AR

D-

CU

RV

ED

PR

OP

ELL

ER

TUB

EA

XIA

L

AX

IAL

FAN

S

VAN

EA

XIA

L

CE

NTR

IFU

GA

L FA

NS

TUB

ULA

R

CE

NTR

IFU

GA

L

SP

EC

IAL

DE

SIG

NS

PO

WE

R R

OO

F V

EN

TILA

TOR

S

AX

IAL

CE

NTR

IFU

GA

L• 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

• Uses same housing configuration as airfoil design.• 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.

• 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.

• Higher pressure characteristics than airfoil, backward-curved, and backward-inclined fans.• Curve may have a break to left of peak pressure and fan should not be operated in this area.• Power rises continually to free delivery.

• 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.

• Simple circular ring, orifice plate, or venturi.• Optimum design is close to blade tips and forms smooth airfoil into wheel.

• Cylindrical tube with close clearance to blade tips.

• Cylindrical tube with close clearance to blade tips.• Guide vanes upstream or downstream from impeller increase pressure capability and efficiency.

• 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.

• Normal housing not used, since air discharges from impeller in full circle.• Usually does not include configuration to recover velocity pressure component.

• 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.

• 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.

• Somewhat more efficient and capable of developing more useful static pressure than propeller fan.• Usually has 4 to 8 blades with airfoil or single- thickness cross section.• Hub usually less than transfer by velocity pressure.

• 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.

• 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.

• 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.

• 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.

R

M

A

B

R

M

Figure 4.2 - Types of Fans

Adapted with permission from 1996 ASHRAE Systems and Equipment Handbook (SI)

12

AMCA 201-02 (R2007)

Figure 4.2 - Types of Fans

Adapted with permission from 1996 ASHRAE Systems and Equipment Handbook (SI)

PERFORMANCE CHARACTERISTICS APPLICATIONSPERFORMANCE CURVES a

• Similar to airfoil fan, except peak efficiency slightly lower.

• Higher pressure characteristics than airfoil and backward- curved fans.• Pressure may drop suddenly at left of peak pressure, but this usually causes no problems.• Power rises continually to free delivery.

• 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.

• High flow rate, but very low-pressure capabilities.• Maximum efficiency reached near free delivery.• Discharge pattern circular and airstream swirls.

• 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.

• 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.

• Usually operated without ductwork; therefore, operates at very low pressure and high volume.• Only static pressure and static efficiency are shown for this fan.

• 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.

• Has straight-through flow.

• Primarily for low-pressure, return air systems in HVAC applications.

• 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.

• 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.

• 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.

• 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.

• Primarily for low-pressure HVAC applications, such as residential furnaces, central station units, and packaged air conditioners.

• 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.

• 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.

• General heating, ventilating, and air-conditioning 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.

• 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.

PR

ES

SU

RE

-PO

WE

R

EFF

ICIE

NC

Y

VOLUME FLOW RATE, Q

10

108

8

6

4

2

0

6

4

2

00 2 4 6 8 10

PR

ES

SU

RE

-PO

WE

R

EFF

ICIE

NC

Y

VOLUME FLOW RATE, Q

10

108

8

6

4

2

0

6

4

2

00 2 4 6 8 10

PR

ES

SU

RE

-PO

WE

R

EFF

ICIE

NC

Y

VOLUME FLOW RATE, Q

10

108

8

6

4

2

0

6

4

2

00 2 4 6 8 10

PR

ES

SU

RE

-PO

WE

R

EFF

ICIE

NC

Y

VOLUME FLOW RATE, Q

10

108

8

6

4

2

0

6

4

2

00 2 4 6 8 10

PR

ES

SU

RE

-PO

WE

R

EFF

ICIE

NC

Y

VOLUME FLOW RATE, Q

10

108

8

6

4

2

0

6

4

2

00 2 4 6 8 10

PR

ES

SU

RE

-PO

WE

R

EFF

ICIE

NC

Y

VOLUME FLOW RATE, Q

10

108

8

6

4

2

0

6

4

2

00 2 4 6 8 10

PR

ES

SU

RE

-PO

WE

R

EFF

ICIE

NC

Y

VOLUME FLOW RATE, Q

10

108

8

6

4

2

0

6

4

2

00 2 4 6 8 10

PR

ES

SU

RE

-PO

WE

R

EFF

ICIE

NC

Y

VOLUME FLOW RATE, Q

10

108

8

6

4

2

0

6

4

2

00 2 4 6 8 10

PR

ES

SU

RE

-PO

WE

R

EFF

ICIE

NC

Y

VOLUME FLOW RATE, Q

10

108

8

6

4

2

0

6

4

2

00 2 4 6 8 10

PR

ES

SU

RE

-PO

WE

R

EFF

ICIE

NC

Y

VOLUME FLOW RATE, Q

10

108

8

6

4

2

0

6

4

2

00 2 4 6 8 10

Ps

Pt

ηt

ηs

wo

• 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.

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.

13

AMCA 201-02 (R2007)

5. Catalog Performance Tables

5.1 Type A: Free inlet, free outlet fans

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.

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 multi-

rating table. Figure 5.1 illustrates such a table for part

of a line of belt-drive propeller fans.

5.2 Ducted fans

There are three types of ducted fans, as described in

Section 3:

1) Type B: Free inlet, ducted outlet

2) Type C: Ducted inlet, free outlet

3) Type D: Ducted inlet, ducted outlet

The performance of fans intended for use with duct

systems is usually published in the form of a "multi-

rating" table. A typical multi-rating table, as illustrated

in Figure 5.2 shows:

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)

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.

Figure 5.1 - Propeller Fan Performance Table

SIZE

(cm)

No. of

Blades

Motor

kWrpm

Peak

kW

AIRFLOW (m3/s) @ STATIC PRESSURE (Pa)

0 31 62 93 124 155 186 217 248

61 3

0.19 862 0.13 2.02 1.58 0.58

0.19 960 0.20 2.25 1.87 0.97

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

69 3

0.19 806 0.20 2.89 2.36 1.05

0.25 883 0.27 3.17 2.68 1.94 0.76

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

84 3

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

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 1.84

SIZE

(in.)

No. of

Blades

Motor

hprpm

Peak

bhp

AIRFLOW (ft3/min) @ STATIC PRESSURE (in. wg)

0 1/8 1/4 3/8 1/2 5/8 3/4 7/8 1

24 3

1/4 862 0.18 4,283 3,350 1,230

1/4 960 0.27 4,770 3,960 2,050

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

27 3

1/4 806 0.27 6,123 4,990 2,230

1/3 883 0.36 6,708 5,675 4,100 1,620

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

33 3

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

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

TYPICAL RATING TABLE FOR A SERIES OF BELT-DRIVEN PROPELLER FANS

Volume

CFM

Outlet

Vel.

(fpm)

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 bhp rpm bhp rpm bhp rpm bhp rpm bhp

3825

4590

5355

6120

6885

500

600

700

800

900

222

236

253

272

292

0.185

0.233

0.292

0.365

0.450

270

284

300

317

0.334

0.400

0.483

0.579

313

327

343

0.519

0.608

0.716

352

366

0.743

0.856 389 1.01 411 1.17

7650

8415

9180

9945

10710

1000

1100

1200

1300

1400

314

338

361

385

409

0.560

0.682

0.826

0.989

1.175

337

358

379

402

425

0.695

0.832

0.988

1.163

1.360

360

378

398

419

441

0.840

0.981

1.149

1.340

1.553

383

399

417

437

457

0.992

1.144

1.314

1.514

1.741

403

419

436

454

473

1.15

1.31

1.49

1.69

1.93

424

438

455

472

489

1.31

1.48

1.68

1.89

2.12

443

458

472

489

506

1.48

1.60

1.86

2.09

2.34

494

507

522

538

2.04

2.25

2.49

2.76

540

554

568

2.67

2.92

3.20

11475

12240

13005

13770

14535

1500

1600

1700

1800

1900

434

458

483

508

1.387

1.626

1.895

2.191

449

473

498

522

547

1.587

1.837

2.115

2.424

2.767

464

488

511

535

559

1.780

2.048

2.346

2.665

3.017

479

501

525

538

571

1.993

2.269

2.570

2.901

3.275

494

515

537

560

584

2.19

2.49

2.80

3.15

3.52

509

529

550

572

595

2.40

2.70

3.03

3.40

3.78

524

543

564

585

606

2.61

2.92

3.26

3.64

4.04

555

572

590

610

630

3.06

3.39

3.73

4.12

4.55

584

600

617

635

654

3.52

3.87

4.24

4.63

5.07

15300

16830

18360

19890

21420

2000

2200

2400

2600

2800

571

621

3.144

4.003

585

633

682

3.403

4.289

5.335

595

644

693

742

791

3.672

4.577

5.632

6.885

8.308

607

654

703

752

801

3.93

4.87

5.96

7.22

8.67

618

665

712

761

810

4.21

5.16

6.28

7.56

9.03

629

675

721

769

818

4.48

5.46

6.61

7.91

9.40

651

695

741

788

834

5.02

6.06

7.24

8.60

10.15

674

715

759

805

852

5.56

6.65

7.90

9.30

10.88

22950

24480

26010

27540

29070

30600

3000

3200

3400

3600

3800

4000

850 10.32 859

908

10.71

12.50

867

916

965

1015

11.09

13.01

15.16

17.52

883

932

981

1030

1072

1129

11.89

13.84

16.03

18.47

21.16

24.11

898

946

995

1044

1093

1142

12.70

14.70

16.92

19.39

22.13

25.16

IMPELLER DIAMETER: 36.5 IN OUTLET AREA: 7.65 SQ FT

TIP SPEED IN FPM: 9.56 × RPM MAXIMUM BHP: 18.3 × (RPM/1000)3

TYPICAL MULTISPEED RATING TABLE FOR A SINGLE WIDTH, SINGLE INLET CENTRIFUGAL FAN

Figure 5.2 - Centrifugal Fan Performance Tables

IMPELLER DIAMETER: 927 mm OUTLET AREA: .71 SQ METERS

TIP SPEED IN m/s: .0485 × RPM MAXIMUM kW: 13.65 × (RPM/1000)3

Volume

m3/s

Outlet

Vel.

(m/s)

62 Pa 93 Pa 124 Pa 155 Pa 186 Pa 217 Pa 246 Pa 310 Pa 373 Pa

rpm kW rpm kW rpm kW rpm kW rpm kW rpm kW rpm kW rpm kW rpm kW

1.81

2.17

2.53

2.89

3.25

2.55

3.06

3.56

4.07

4.58

222

236

253

272

292

0.14

0.17

0.22

0.27

0.34

270

284

300

317

0.25

0.30

0.36

0.43

313

327

343

0.39

0.45

0.53

352

366

0.55

0.64 389 0.75 411 0.87

3.61

3.97

4.33

4.69

5.06

5.08

5.59

6.10

6.61

7.13

314

338

361

385

409

0.42

0.51

0.62

0.74

0.88

337

358

379

402

426

0.52

0.62

0.74

0.87

1.01

360

378

398

419

441

0.63

0.73

0.86

1.00

1.16

382

399

417

437

457

0.74

0.85

0.98

1.13

1.30

403

419

436

454

473

0.86

0.98

1.11

1.26

1.44

424

438

455

472

489

0.98

1.10

1.25

1.41

1.58

443

458

472

489

506

1.10

1.19

1.39

1.56

1.74

494

507

522

538

1.52

1.68

1.86

2.06

540

554

568

1.99

2.18

2.39

5.42

5.78

6.14

6.50

6.86

7.63

8.14

8.65

9.15

9.66

434

458

483

508

1.03

1.21

1.41

1.63

449

473

498

522

547

1.18

1.37

1.58

1.81

2.06

464

488

511

535

559

1.33

1.53

1.75

1.99

2.25

479

501

525

538

571

1.49

1.69

1.92

2.16

2.44

494

515

537

560

584

1.63

1.86

2.09

2.35

2.62

509

529

550

572

595

1.79

2.01

2.26

2.54

2.82

524

543

564

585

606

1.95

2.18

2.43

2.71

3.01

555

572

590

610

630

2.28

2.53

2.78

3.07

3.39

584

600

617

635

654

2.62

2.89

3.16

3.45

3.78

7.22

7.94

8.67

9.39

10.11

10.17

11.18

12.21

13.23

14.24

571

621

2.34

2.99

585

633

682

2.54

3.20

3.98

595

644

693

742

791

2.74

3.41

4.20

5.13

6.20

607

654

703

752

801

2.93

3.63

4.44

5.38

6.47

616

665

712

761

810

3.14

3.85

4.68

5.64

6.73

629

675

721

769

818

3.34

4.07

4.93

5.90

7.01

651

695

741

788

834

3.74

4.52

5.40

6.41

7.57

674

715

759

805

852

4.15

4.96

5.89

6.94

8.11

10.83

11.55

12.28

13.00

13.72

14.44

15.25

16.27

17.30

18.31

19.32

20.34

850 7.70 859

908

7.99

9.40

867

916

965

1015

8.27

9.70

11.30

13.06

883

932

981

1030

1072

1129

8.87

10.32

11.95

13.77

15.78

17.98

898

946

995

1044

1093

1142

9.47

10.96

12.62

14.46

16.50

18.76

TYPICAL MULTISPEED RATING TABLE FOR A SINGLE WIDTH, SINGLE INLET CENTRIFUGAL FAN

AMCA 201-02 (R2007)

14

222

236

253

272

292

.185

.233

.292

.365

.450

270

284

300

317

.334

.400

.483

.579

313

327

343

.51

9.6

08

.71

6352

366

.743

.856

389

1.0

1411

1.1

7

314

338

361

335

409

.560

.682

.826

.988

1.1

75

337

358

379

482

426

.695

.822

.988

1.1

63

1.3

60

360

378

398

419

441

.84

0.9

81

1.1

49

1.3

40

1.5

53

332

399

417

437

457

.992

1.1

44

1.3

14

1.5

14

1.7

41

403

419

436

454

473

1.1

51.3

11.4

91.6

91.9

3

424

438

455

472

489

1.3

11.4

81.5

81.8

92.1

2

443

458

472

489

506

1.4

81.6

01.8

62.0

92.3

4

494

507

522

538

2.0

42.2

52.4

92.7

6

540

554

568

2.6

72.9

23.2

8584

598

3.3

73.6

6

434

456

482

508

1.3

87

1.6

26

2.1

9

449

473

493

522

547

1.5

87

1.8

37

2.1

15

2.4

24

2.7

67

464

488

511

535

559

1.7

82.0

48

2.3

46

2.6

65

3.0

17

479

501

525

538

571

1.9

95

2.2

69

2.5

70

2.9

01

3.2

76

494

515

537

560

584

2.1

92.4

92.8

03.1

53.5

2

509

529

550

572

595

2.4

02.7

03.0

33.4

0

524

543

564

585

606

2.6

12.9

23.2

63.8

44.0

4

555

572

590

610

630

3.0

63.4

93.7

34.1

24.5

5

584

600

617

635

654

3.5

23.8

74.2

44.6

35.0

7

612

627

643

661

678

3.9

94.3

64.7

65.1

85.6

3

571

629

3.7

44

4.0

03

584

633

682

3.4

03

4.2

89

5.3

35

596

644

693

742

791

4.5

77

5.6

32

6.8

85

8.3

08

607

654

703

752

801

3.9

34.8

75.7

67.2

28.6

7

618

665

712

761

810

4.2

15.1

66.2

87.5

69.0

3

629

675

721

769

818

4.4

85.4

66.8

17.9

18.4

8

651

695

741

788

834

5.0

26.0

67.2

48.6

010.1

5

674

715

759

852

5.5

66.6

57.9

09.3

010.8

8

696

736

778

822

867

6.1

17.2

4

10.0

211

.65

850

10.3

2859

908

10.7

112.6

0867

916

965

10

15

11.0

913.0

115.1

617.5

2

883

932

981

1030

1079

1129

11.8

913.8

416.0

318.4

721.1

624.1

1

898

946

995

1044

1093

1142

12.7

014.7

016.9

219.3

922.1

325.1

6

914

960

10

09

1057

1106

1155

13.4

815.5

617.8

320.3

523.1

226.1

8

RECOMMENDEDSELECTION RANGE810 RPM585 RPM

490 RPM

390 RPM

PR

ES

SU

RE

IN IN

. WG

BR

AK

E H

OR

SE

PO

WE

R

VO

LUM

EC

FMO

UTL

ET

VE

LOC

ITY

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

3825

4590

5355

6120

6885

7650

8415

9180

9945

1071

0

1147

512

240

1300

513

770

1453

5

1530

016

830

1836

019

890

2142

0

2295

024

480

2601

027

540

2907

030

600

CFM

1/4”

SP

3/8”

SP

1/2”

SP

5/8”

SP

3/4”

SP

7/8”

SP

1” S

P1-

1/4”

SP

1-1/

2” S

P1-

3/4”

SP

RP

MB

HP

RP

MB

HP

RP

MB

HP

RP

MB

HP

RP

MB

HP

RP

MB

HP

RP

MB

HP

RP

MB

HP

RP

MB

HP

RP

MB

HP

AMCA 201-02 (R2007)

15

Figure 5.3 - Typical Fan Performance Table Showing Relationship to a Family

of Constant Speed Performance Curves

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.

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.

Fan performance adjustment factors for airstream

accessories are normally available from either the fan

catalog or the fan manufacturer.

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.

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.

6.2 Component losses

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.

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.

AIRFLOW

PR

ES

SU

RE

SELECTION NOT USUALLY

RECOMMENDED IN THIS RANGE

SELECTION

NOT USUALLY

RECOMMENDED

IN THIS RANGE

RECOMMENDED

SELECTION RANGE

PR

ESSU

RE

DUCT SYSTEM CURVE

DU

CT S

YSTEM

CU

RVE

Figure 5.4 - Recommended Performance Range of a Typical Centrifugal Fan

AMCA 201-02 (R2007)

16

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.

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:

Pc/P = (Qc/Q)2

A more through discussion of duct system pressure

losses can be found in AMCA Publication 200 AirSystems.

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.

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 Qwill 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.

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.

SYSTE

M B

SYSTEM A

SYSTEM C

PE

RC

EN

T O

F S

YS

TE

M R

ES

ISTA

NC

E

PERCENT OF SYSTEM AIRFLOW

0

20

40

60

80

100

120

140

160

180

200

0

20 40 60 80 100 120 140 160 180 200

SYSTEMDESIGN POINT

Figure 6.1 - System Curves

AMCA 201-02 (R2007)

17

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 SEFhave been accurately determined, then the fan will

deliver the designated airflow when installed in the

system.

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.

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

is now at Point 3 (the intersection of the fan curve and

the new System C), with the airflow at approximately

120% of Q.

6.5 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 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.

200

0

20

40

60

80

100

120

140

160

180

200

40 60 80 100 120 140 160 180 200

FAN CURVE

SYSTEM B

SY

STE

M A

SYSTEM C

SYSTEMDESIGN POINT

1

2

3


PE

RC

EN

T O

F S

YS

TE

M R

ES

ISTA

NC

E

Figure 6.2 - Interaction of System Curves and Fan Curve

AMCA 201-02 (R2007)

18


PE

RC

EN

T O

F P

OW

ER

PE

RC

EN

T O

F S

YS

TE

M R

ES

ISTA

NC

E

0

0

20

40

60

80

100

120

140

160

20 40 60 80 100

100

133

50

120 140

110%

160 180 200

H (AT 1.1N)PRESSURES (AT 1.1N) D

UC

T S

YS

TE

M A

PRESSURES (AT N)

H (AT N)1

2

Figure 6.3 - Effect of 10% increase in Fan Speed

AMCA 201-02 (R2007)

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.

Qc = Q × (Nc/N)

Ptc = Pt × (Nc/N)2

Psc = Ps × (Nc/N)2

Pvc = Pv × (Nc/N)2

Hc = H × (Nc/N)3

6.6 Effect of density on system resistance

The resistance of a duct system is dependent upon

the density of the air flowing through the system. An

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.

6.6.1 Fan Laws - effect of change in density - (fan

size and speed remaining constant)

When the speed of the fan does not change, Nc = Nand, therefore, (Nc/N) = 1. The fan size is also fixed,

Dc = D and therefore (Dc/D) = 1.

Qc = Q

Ptc = Pt × (ρc/ρ)

Psc = Ps × (ρc/ρ)

Pvc = Pv × (ρc/ρ)

Hc = H × (ρc/ρ)

19

0

0

0 20 40 60 80 100 120 140 160 180 200

20

40

60

80

100

20

40

60

80

100


PE

RC

EN

T O

F P

OW

ER

PE

RC

EN

T O

F S

YS

TE

MR

ES

ISTA

NC

E A

ND

FA

N P

RE

SS

UR

E

POWER @ DENSITY ρ

FAN PRESSURE CURVE@ DENSITY ρ/2

FAN PRESSURE CURVE@ DENSITY ρ SYSTEM A

@ DENSITY ρFAN INLET

SYSTEM A@ DENSITY ρ/2

FAN INLET

POWER @ DENSITY ρ/2

Figure 6.4 - Density Effect

AMCA 201-02 (R2007)

20

CALCULATED SYSTEM CURVE

PEAK FAN PRESSURE

FAN PRESSURE

CURVE

DESIGN AIRFLOW

DE

SIG

N R

ES

ISTA

NC

E

1

Figure 6.5 - Fan/System Curve at Design Point

AMCA 201-02 (R2007)

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.

21

FAN PRESSURECURVE

CURVE B:ACTUAL SYSTEM

CURVE A:CALCULATED SYSTEM

CURVE CACTUAL SYSTEM

PEAK FANPRESSURE

ACTUAL SYSTEM RESISTANCEMORE THAN DESIGN

ACTUAL SYSTEMLESS THANDESIGN

DESIGN AIRFLOW

DE

SIG

N R

ES

ISTA

NC

E

5

1

2

4

3

Figure 6.6 - Fan/System Curve Not at Design Point

AMCA 201-02 (R2007)

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 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.

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.

22

AMCA 201-02 (R2007)

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 to prevent deficient

performance

• 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 FieldPerformance Measurement of Fan Systems;

for more precise measurement see AMCA

Standard 803 Industrial Process/PowerGeneration Fans: Site Performance TestStandard. 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 SEFequal 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.

23

AMCA 201-02 (R2007)

7. System Effect Factor (SEF)

A System Effect Factor is a value that accounts for

the effect of conditions adversely influencing fan

performance when installed in the air system.

7.1 System Effect Curves

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.

24

AIRFLOWDEFICIENCY

SYSTEMEFFECT ATACTUAL AIRFLOW

FAN CATALOGPRESSURECURVE

SYSTEM EFFECT LOSSAT DESIGN AIRFLOW

CURVE ACALCULATED SYSTEMWITH NO ALLOWANCEFOR SYSTEM EFFECT

CURVE BACTUAL SYSTEMWITH SYSTEM EFFECT

DESIGN AIRFLOW

DE

SIG

N R

ES

ISTA

NC

E

1

2

4

3

Figure 6.7 - Deficient Fan/System Performance - System Effect Ignored

AIR VELOCITY, (m/s)

SY

ST

EM

EF

FE

CT

FA

CT

OR

PR

ES

SU

RE

, P

a

(Air Density = 1.2 kg/m3)

1000

900

800

700

600

500

400

300

200

100

90

80

70

60

50

40

30

20

2.5 3 4 5 6 7 8 9 10 20 30

X

W

V

U

T

S

R

Q

PONMLKJIHGF

Figure 7.1 - System Effect Curves (SI)

AMCA 201-02 (R2007)

25

AIR VELOCITY, ft/min × 100

SY

ST

EM

EF

FE

CT

FA

CT

OR

- P

RE

SS

UR

E,

in.

wg

50.1

0.15

0.2

0.25

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1.0

1.5

2.0

2.5

3.0

4.0

5.0

6 7 8 9 10 15 20 25 30 40 50 60

FG H I J K L M N O

P

Q

R

S

T

U

V

W

X

(Air Density = 0.075 lbm/ft3)

Figure 7.1 - System Effect Curves (I-P)

AMCA 201-02 (R2007)

26

Table 7.1 - System Effect Coefficients

Curve in Dynamic Pressure

Figure 7.1 Loss Coefficient C

F 16.00

G 14.20

H 12.70

I 11.40

J 9.50

K 7.90

L 6.40

M 4.50

N 3.20

O 2.50

P 1.90

Q 1.50

R 1.20

S 0.75

T 0.50

U 0.40

V 0.25

W 0.17

X 0.10

SI

I-P

SEF C V= ⎛⎝⎜

⎞⎠⎟1097

2

ρ

SEF C V= ⎛⎝⎜

⎞⎠⎟1 414

2

.ρ

AMCA 201-02 (R2007)

27

DESIGN AIRFLOW

AC

TU

AL S

YS

TE

M R

ES

ISTA

NC

E

AC

TU

AL P

OW

ER

RE

QU

IRE

D

SEF

FAN POWER

FAN PRESSURE

ACTUAL SYSTEM W/ SEF

CALCULATEDSYSTEM W/NOALLOWANCE

FOR SEF

Figure 7.2 - Effect of System on Fan Selection

AMCA 201-02 (R2007)

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:

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.

SEF SEF dd2 1

2

1

=⎛

⎝⎜

⎞

⎠⎟

28

OUTLET AREA

BLAST AREA

CENTRIFUGAL FAN

AXIAL FAN

CUTOFF

DISCHARGE DUCT

25%

50%

75%

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. If the duct is rectangular with side dimensions aand b, the equivalent duct diameter is equal to (4ab/π)0.5.

Figure 8.1 - Fan Outlet Velocity Profiles

AMCA 201-02 (R2007)

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. Outlet System Effect Factors

8.1 Outlet ducts

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

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.

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.

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).

29

AXIAL FAN


Figure 8.2 - System Effect Curves for Outlet Ducts - Axial Fans

Tubeaxial Fan

Vaneaxial Fan

No Duct

12%

Effective

Duct

25%

Effective

Duct

50 %

Effective

Duct

100%

Effective

Duct

--- --- --- --- ---

U V W --- ---

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

Determine SEF by using Figure 7.1

AMCA 201-02 (R2007)

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

30

OUTLET AREA

BLAST AREA

CENTRIFUGAL FAN

CUTOFF

DISCHARGE DUCT


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.

Figure 8.3 - System Effect Curves for Outlet Ducts - Centrifugal Fans

No Duct12%

Effective Duct

25%

Effective Duct

50%

Effective Duct

100%

Effective Duct

Pressure

Recovery0% 50% 80% 90% 100%

Blast AreaOutlet Area System Effect Curve

0.4

0.5

0.6

0.7

0.8

0.9

1.0

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

—

—

W

W

W-X

—

—

—

—

—

—

—

—

—

—

—

Determine SEF by using Figure 7.1

AMCA 201-02 (R2007)

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.

31

TUBEAXIAL FAN SHOWN

VANEAXIAL FAN SHOWN

% EFFECTIVEDUCT LENGTH

% EFFECTIVEDUCT LENGTH

Determine SEF by using Figures 7.1 and 8.1

Figure 8.4 - System Effect Curves for Outlet Duct Elbows - Axial Fans

Tubeaxial Fan

Vaneaxial Fan

Vaneaxial Fan

90° Elbow No Duct

12%

Effective

Duct

25%

Effective

Duct

50 %

Effective

Duct

100%

Effective

Duct

2 & 4 Pc --- --- --- --- ---

2 Pc U U-V V W ---

4 Pc W --- --- --- ---

AMCA 201-02 (R2007)

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 industry-

accepted value.

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)

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.

8.3.1 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).

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.

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.

32

POSITION C

POSITION D

POSITION B

POSITION A

SWSI CENTRIFUGAL FAN SHOWN

INLET

% EFFECTIVE

DUCT LENGTH

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

AMCA 201-02 (R2007)

33

Blast AreaOutlet Area

Outlet

Elbow

Position

No Outlet

Duct

12%

Effective

Duct

25%

Effective

Duct

50%

Effective

Duct

100%

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

NO

Sys

tem

Effe

ct F

acto

r

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 = ΔP × 1.00

ELBOW POSITION B = ΔP × 1.25

ELBOW POSITION C = ΔP × 1.00

ELBOW POSITION D = ΔP × 0.85

Figure 8.5 - Outlet Elbows on SWSI Centrifugal Fans

AMCA 201-02 (R2007)

34

PARALLEL-BLADE DAMPERILLUSTRATING DIVERTED AIRFLOW

OPPOSED-BLADE DAMPERILLUSTRATING NON-DIVERTEDAIRFLOW

Figure 8.6 - Parallel Blade vs. Opposed Blade Damper

AMCA 201-02 (R2007)

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 non-

uniform 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

a large plenum or to free space a parallel blade

damper may be satisfactory.

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 non-

uniform 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.

35

VOLUME CONTROL DAMPER

Figure 8.7 - Pressure Drop Multipliers for Volume Control Dampers on a Fan Discharge

BLAST AREA PRESSURE DROP

OUTLET AREA 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

AMCA 201-02 (R2007)

36

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

AMCA 201-02 (R2007)

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.

37

Figure 9.1 Typical Inlet Connections for Centrifugal and Axial Fans

CONVERGING TAPERED ENTRYINTO FAN OR DUCT SYSTEM

FLANGED ENTRY INTOFAN OR DUCT SYTEM

IDEAL SMOOTH ENTRY TODUCT ON A DUCT SYSTEM

a.BELL MOUTH INLET PRODUCESFULL FLOW INTO FAN

b.VENA CONTRACTA AT INLETREDUCES EFFECTIVE FAN INLET AREA

c.

e.d.

AMCA 201-02 (R2007)

9. Inlet System Effect Factors

Fan performance can be greatly affected by non-

uniform 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.

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).

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 venacontracta. 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

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.

ANSI/AMCA 210 limits an inlet duct to a cross-

sectional 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.

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.

38

DUCT LENGTH

DUCT LENGTH

VANEAXIAL FAN SHOWN

TUBEAXIAL 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

AMCA 201-02 (R2007)

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 hub-

to-tip (H/T) ratio and blade solidity. The number of

blades did not have a significant affect on the inlet

elbow SEF.

9.2.2 Centrifugal fans - inlet duct elbows. Non-

uniform 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 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

39

40

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.

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).

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.

Figure 9.3A - Non-Uniform Airflow Into a Fan

Inlet Induced by a 90°, 3-Piece Section Elbow--

No Turning Vanes

Figure 9.3B - Non-Uniform Airflow Induced Into

Fan Inlet by a Rectangular Inlet Duct

LENGTHOF DUCT

D

R

AMCA 201-02 (R2007)

41

AMCA 201-02 (R2007)

LENGTHOF DUCT

R

D

+

LENGTHOF DUCT

D

R

+

LENGTHOF DUCT D

R

+

DETERMINE SEF BY USING FIGURE 7.1

Figure 9.4 - System Effect Curves for Various Mitered Elbows without Turing Vanes

Figure 9.4A - Two Piece Mitered 90° Round Section Elbow - Not Vaned

Figure 9.4B - Three Piece Mitered 90° Round Section Elbow - Not Vaned

Figure 9.4C - Four or More Piece Mitered 90° Round Section Elbow - Not Vaned

SYSTEM EFFECT CURVES

R/D NO 2D 5D

DUCT DUCT DUCT

— N P R-S


R/D NO 2D 5D

DUCT 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/D NO 2D 5D

DUCT 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 = 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

H

H

+ R

LENGTHOF DUCT

H

H

+ R

LENGTHOF DUCT

H

H

+

LENGTHOF DUCT

R


R/D NO 2D 5D

DUCT 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


R/D NO 2D 5D

DUCT DUCT DUCT

0.5 S T-U V

1.0 T U-V W

2.0 V V-W W-X


R/D NO 2D 5D

DUCT DUCT DUCT

0.5 S T-U V

1.0 T U-V W

2.0 V V-W W-X

Figure 9.5B - Square Elbow with Inlet Transition - 3 Long Turning Vanes

Figure 9.5A - Square Elbow with Inlet Transition - No Turning Vanes

Figure 9.5C - Square Elbow with Inlet Transition - Short Turning Vanes

AMCA 201-02 (R2007)

42

IMPELLER

ROTATION

COUNTER-ROTATING SWIRL

Figure 9.7 - Example of a Forced Inlet Vortex

Figure 9.8 - Inlet Duct Connections Causing Inlet Spin

IMPELLERROTATION

IMPELLERROTATION

PRE-ROTATING SWIRL COUNTER-ROTATING SWIRL

AMCA 201-02 (R2007)

43

Figure 9.6 - Improved Flow Conditions with a Special Designed Inlet Box

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

airflow entering a duct elbow with turning vanes will

leave the duct elbow with non-uniform airflow.

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.

TURNINGVANES

TURNINGVANES

TURNINGVANES

CORRECTED PRE-ROTATING SWIRL

CORRECTED COUNTER-ROTATING SWIRL

IMPELLERROTATION

IMPELLERROTATION

Figure 9.9 - Corrections for Inlet Spin

AMCA 201-02 (R2007)

44

DUCT

0.075D

0.075D

D

0.45D

2D

D

DUCTDUCT

Figure 9.10B - ISO 5801 Star Straightener

Figure 9.10A - ANSI/AMCA Standard 210 Egg-Crate Straightener

Figure 9.10 - Test Standard Airflow Straighteners

AMCA 201-02 (R2007)

45

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

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.

2LL L

INLE

TD

IA.

Figure 9.11C - Centrifugal Fan Near Wall(s) Figure 9.11D - DWDI Fan Near Wall on One Side

Figure 9.11A - Fans and Plenum Figure 9.11B - Axial Fan Near Wall

EQUAL

DIAMETEROF INLET

EQUAL L

LL L L

DWDI SWSI

L - DISTANCE

INLET TO WALL

0.75 x DIA. OF INLET




V-W

U

T

S

X

V-W

V-W

U

For Figures 9.11A, B & C


For Figures 9.11D


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

AMCA 201-02 (R2007)

46

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.

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.

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

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.

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.

1 DIA.MIN

Figure 9.13 - Parallel Installation of Axial Flow Fans

Figure 9.12 - Fan in Plenum with Non-Symmetrical Inlet

SPLITTER SHEET

Figure 9.12A - Enclosure Inlet Not Symmetrical with Fan Inlet. Pre-

Rotational 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

AMCA 201-02 (R2007)

47

INLET PLANE

FREE INLET AREA PLANE - FAN WITH INLET COLLAR

FREE INLET AREA PLANE - FAN WITHOUT INLET COLLAR

INSIDE DIAMETER

INLET COLLAR

POINT OF TANGENTWITH FAN HOUSING SIDEAND INLET CONE RADIUS

INLET PLANE

DIAMETER

OF TANGENT

Figure 9.14 - System Effect Curves for Inlet Obstructions

(Table based on Fans and Fan Systems, Thompson & Trickler, Chem Eng MAR83, p. 60)

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

diameterAt 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

AMCA 201-02 (R2007)

48

Table for Figure 9.14

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 TypeVariable

Inlet Vanes

Figure 10.1 - Common Terminology for Centrifugal Fan Appurtenances

AMCA 201-02 (R2007)

49

AMCA 201-02 (R2007)

10.1 Bearing and supports in fan inlet

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.

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.

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.

10.3 Belt tube in axial fan inlet or outlet

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).

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.

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.

50

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).

When variable inlet vanes are supplied by the fan

manufacturer, the performance should include the

effects of the variable inlet vane unit.

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.

20

0 20 40 60 80 100 120

40

60

80

100

120

PERCENT OF WIDE OPEN VOLUME

PE

RC

EN

T O

F S

HU

T-O

FF

PR

ES

SU

RE

75% OPEN

75% OPEN

75% OPEN

FAN PERFORMANCEW/OUT VARIABLE INLET VANES

VARIABLE INLET VANES100% OPEN

CONE TYPE

VARIABLE INLET

VANES

CYLINDRICAL TYPE

VARIABLE INLET

VANES

Figure 10.2 - Typical Variable Inlet Vanes for a Backward Inclined Fan

VANE TYPE SYSTEM EFFECT CURVE

(100% Open)

a) Cone type, integral “Q” or “R”

b) Cylindrical type “S”

Determine SEF by calculating inlet velocity and using

Figure 7.1

AMCA 201-02 (R2007)

51

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

AMCA 201-02 (R2007)

52

AMCA 201-02 (R2007)

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.

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.

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 pressure-

volume 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).

53

100%

100%

200%

PERCENT OF FAN AIRFLOW

PE

RC

EN

T O

F F

AN

STA

TIC

PR

ES

SU

RE

SYSTEMRESISTANCE

SERIES FANCOMBINEDPRESSURE CURVE

SINGLE FANPRESSURE CURVE

Figure B.1 - Typical Characteristic Curve of Two Fans Operating in Series

AMCA 201-02 (R2007)

54

55

AMCA 201-02 (R2007)

PERCENT OF FAN AIRFLOW

PE

RC

EN

T O

F F

AN

STA

TIC

PR

ES

SU

RE

200

100

FAN OPERATION NOTRECOMMENDED IN THISRANGE

PARALLEL FANS - FAN PRESSURE ATCOMBINED VOLUME

SINGLE FAN -PRESSURECURVE

UN

STA

BL

E S

YS

TE

M

STA

BL

E S

YS

TE

M

Figure B.2 - Parallel Fan Operation

AILERON

Figure B.3 - Aileron Control

56

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 AMCA99-0066).

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 (perAMCA 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 99-0066).

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).

C.1.8 Temperature. The dry-bulb temperature (td) isthe 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 dry-

bulb 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 (perAMCA 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.

AMCA 201-02 (R2007)

57

AMCA 201-02 (R2007)

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:

Pta = 0

Pt = Pt2 - Pt1

Pt1 = Pta = 0

Pt = Pt2

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

PLANE 2PLANE 1

Pt2

Pt = Pt2

Figure C.1 - Fan Total Pressure for Installation Type B: Free Inlet, Ducted Outlet

58

AMCA 201-02 (R2007)

PLANE 2PLANE 1

Pt1

Pt = Pv2 - Pt1

Figure C.2 - Fan Total Pressure for Installation Type C: Ducted Inlet, Free Outlet

Figure C.3 - Fan Total Pressure for Installation Type D: Ducted Inlet, Ducted Outlet

PLANE 2PLANE 1

Pt2Pt1

Pt = Pt2 - Pt1

Pt

59

AMCA 201-02 (R2007)

PLANE 2PLANE 1

Pv2

Pv = Pv2

Figure C.4 - Fan Velocity Pressure for Installation Type B: Free Inlet, Ducted Outlet

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

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

60

AMCA 201-02 (R2007)

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 pressure-

volume 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).

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 (perAMCA 99-0066).

Figure C.5 - Fan Static Pressure for Installation Type D: Ducted Inlet, Ducted Outlet

PLANE 2PLANE 1

Ps2Pv1Ps1

Ps = Ps2 - Ps1 - Pv1 (algebraic)

61

AMCA 201-02 (R2007)

HOUSING

DIVERTER

CENTER PLATE

SIDE SHEET

CUT OFF

BEARINGSUPPORT

INLET COLLAR

INLET

BLADE

BACKPLATE

IMPELLER

RIM

CUT OFF

BLAST AREADISCHARGE

OUTLET AREA

SCROLL

FRAME

Figure C.6 - Terminology for Centrifugal Fan Components

62

AMCA 201-02 (R2007)

BELT TUBE

CASING

BEARING CASING

BLADE

HUB

IMPELLER

GUIDE VANE

Figure C.7C - Vaneaxial Fan-Belt Drive

Figure C.7B - Tubeaxial Fan-Direct Drive (Impeller Downstream)

DIFFUSERBLADE

HUB

IMPELLER

INLET BELL

CASING

MOTOR

Figure C.7A - Tubular Centrifugal Fan-Direct Drive

INLET

BACKPLATERIM

HUB

IMPELLER

BLADEGUIDE VANE

MOTOR

CASING

Figure C.7 - Terminology for Axial and Tubular Centrifugal Fans

63

AMCA 201-02 (R2007)

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).

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.

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

C-D Duct friction at Q =

1.42 m3/s (3000 cfm). 747.00 Pa (duct design) 3.0 in. wg

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

ATMOSPHERIC PRESSURE

0

1

2

3

4

0

249

498

747

996

Pt

Ps

Pv

Pv = 124 Pa (0.5 in.wg)

FRICTION 747 Pa (3.0 in.wg)AT 1.42 m3/s (3000 cfm)

A B C D

(I-P) in.wg

(SI) Pa

124 Pa(0.5 in.wg)

Figure D.1 - Pressure Gradients - Fan as Tested

64

AMCA 201-02 (R2007)

SI I-P

C-D Outlet duct on fan as tested 0.00 Pa (no SEF) 0.0 in. wg

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 0.0 in. wg

E Contraction loss - plenum

to duct 49.80 Pa (part of duct system) 0.2 in. wg

E Ps energy required to

create velocity at E 124.50 Pa (part of duct system) 0.5 in. wg

E-F Duct friction at Q =

1.42 m3/s (3000 cfm) 747.00 Pa 3.0 in. wg


Solution:

Select a fan for Q = 1.42 m3/s (3000 cfm) and Ps = 921.30 Pa (3.7 in. wg)


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).

65

AMCA 201-02 (R2007)



5

4

3

2

1

0

1245

996

747

498

249

0

Pt

Ps

Pv

Pv = 124 Pa (0.5in.wg)

1046 Pa (3.7 in.wg)

922 Pa (3.7 in.wg)

747 Pa (3.0 in.wg)

922 Pa (3.7 in.wg)

A B C D EF

(I-P) in.wg

(SI) Pa

124 Pa(0.5 in.wg)

NEGLIGIBLELOSS

2.5 DIA.

Figure D.2 - Pressure Gradients - Plenum Effect

66

AMCA 201-02 (R2007)

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)

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.

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 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".

The SEF for fans without outlet duct was obtained as

follows:

GIVEN:

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)

Fan Blast AreaOutlet Area

= 0 6.

SI I-P

B-C SEF 149.40 Pa 0.6 in. wg

(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 0.0 in. wg

D contraction loss - plenum

to duct 49.80 Pa (part of duct system) 0.2 in. wg

D Ps energy required to

create velocity at D 124.50 Pa (part of duct system) 0.5 in. wg

D-E duct friction at Q =

1.42 m3/s (3000 cfm) 747.00 Pa (duct design) 3.0 in. wg


Solution:

Select a fan for 1.42 m3/s (3000 cfm) Q and 1070.70 Pa (4.3 in. wg) Ps.


67

AMCA 201-02 (R2007)



5

4

3

2

1

0

1245

996

747

498

249

0

Pt

Ps

Pv

Pv = 124 Pa (0.5 in.wg)

922 Pa (3.7 in.wg)

872 Pa (3.5 in.wg)

747 Pa (3.0 in.wg)

149 Pa (0.6 in.wg) SEF

A B C D E

(I-P) in.wg

(SI) Pa

124 Pa (0.5 in.wg)

Figure D.3 - Pressure Gradients - Abrupt Expansion at Fan Outlet

68

AMCA 201-02 (R2007)

SI I-P

A 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


D.4 Example of fan used to exhaust with

obstruction in inlet, inlet elbow, inlet duct,

free outlet

This example is an exhaust system. Note the entry

loss at point A. An inlet bell will reduce this loss.

On the suction side of the fan, Ps will be negative, but

Pv is always positive.

Fan Pv = 124.5 Pa (0.5 in. wg)

Three SEFs are shown in this example:

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).

2) System Effect Curve U (see Figure 9.14 for a

bearing in the fan inlet which obstructs 10% of the

inlet).

3) System Effect Curve R (from Figure 8.3 for a fan

discharging to atmosphere with no outlet duct).

69

AMCA 201-02 (R2007)



FAN INLET

-5

-4

-3

-2

-1

0

+1

-1245

-996

-747

-498

-249

0

+249

Pv

Pt

Ps

149 Pa (0.6 in.wg)ELBOW SEF

50 Pa (0.2 in.wg)OBSTRUCTION SEF

149 Pa (0.6 in.wg)

REQUIRED

149 Pa (0.6 in.wg)

ABRUPTDISCHARGE SEF P

v = 124 Pa (0.5 in.wg)

100 Pa (0.4 in.wg)

-847 Pa (-3.4 in.wg)

-996 Pa (4.0 in.wg)

-971 Pa (3.9 in.wg)

-1121 Pa (4.5 in.wg)

-1171 Pa (4.7 in.wg)

224 Pa (0.9 in.wg)

C D E

(I-P) in.wg

(SI) Pa

BA

Figure D.4 - Pressure Gradients - Exhaust System

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 Movementand 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 WestUniversity Drive, Arlington Heights, IL, 60004-1893 U.S.A., 1995.

3. AMCA Publication 202-98, Troubleshooting, Air Movement and Control Association International, Inc., 30 WestUniversity Drive, Arlington Heights, IL, 60004-1893 U.S.A., 1997.

4. ASHRAE Handbook, HVAC Systems and Equipment, 1996, The American Society of Heating, Refrigeratingand Air Conditioning Engineers, Inc., 1791 Tullie Circle N.E., Atlanta, GA, 30329 U.S.A., 1996, (Chapter 18Fans).

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, TheAmerican 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 EngineeringConference, Air Movement and Control Association International, Inc., 30 West University Drive, ArlingtonHeights, 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 FanPerformance, ASHRAE RP-139 Report, The American Society of Heating, Refrigerating and Air ConditioningEngineers, 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 Societyof 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 EngineeringConference, Air Movement and Control Association International, Inc., 30 West University Drive, ArlingtonHeights, 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, AMCAEngineering 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, AMCAEngineering 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 KemperMeadow 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, ChemicalEngineering, March 21, 1983, pp. 48-63

16. AABC National Standards, Chapter 8, Volume Measurements, Associated Air Balance Council, 1518 K StreetNW, Suite 503, Washington, DC 20005 U.S.A.

AMCA 201-02 (R2007)

70

AIR MOVEMENT AND CONTROLASSOCIATION INTERNATIONAL, INC.

30 West University DriveArlington Heights, IL 60004-1893 U.S.A.

E-Mail : [email protected] Web: www.amca.orgTel: (847) 394-0150 Fax: (847) 253-0088

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.

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