AMCA 201-02 (R2011) Fans and Systems

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PUBLICATION

AMCA 201-02 (R2011) Fans and Systems

Air Movement and Control Association International AMCA Corporate Headquarters 30 W. University Drive, Arlington Heights, IL 60004-1893, USA [email protected] n Ph: +1-847-394-0150 n www.amca.org © 2017 Air Movement & Control Association International

Fans and Systems

Air Movement and Control Association International, Inc. 30 West University Drive Arlington Heights, IL 60004-1893

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AMCA PUBLICATION 201-02 (R2011)

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© 2011 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.

ANSI/AMCA Standard 210 Laboratory Methods of Testing Fans for Aerodynamic Performance Rating, provides a basis for accurately rating the performance of fans when tested under standardized laboratory conditions. The actual performance of a fan when installed in an air moving system will sometimes be different from the fan performance as measured in the laboratory. The difference in performance between the laboratory and the field installation can sometimes be attributed to the interaction of the fan and the duct system, i.e., duct system design can diminish the usable output of the fan. AMCA Publication 201 Fans and Systems, introduced the concept of System Effect Factor to the air moving industry. The System Effect Factor quantifies the duct system design effect on performance. The System Effect Factor 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 major revision 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 light of available information and accepted industry practices. However, AMCA International does not guarantee, certify or assure the safety or performance of any products, components or systems tested, designed, installed or operated in accordance with AMCA International standards or that any tests conducted under its standards will be non-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 regarding its standards, certification programs, or interpretations thereof. For information on procedures for submitting and handling complaints, write to: Air Movement and Control Association International 30 West University Drive Arlington Heights, IL 60004-1893 U.S.A. or AMCA International, Incorporated c/o Federation of Environmental Trade Associations 2 Waltham Court, Milley Lane, Hare Hatch Reading, Berkshire RG10 9TH United Kingdom

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Forward

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.

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Related AMCA Standards and Publications

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

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TABLE OF CONTENTS

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

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

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

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Annex A.

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AMCA 201-02 (R2011)

Fans and Systems

values inconsistent with the values presented.

1. Introduction

Mechanical design of the fan is not within the scope of this publication.

ANSI/AMCA 210 Laboratory Methods of Testing Fans For Aerodynamic Performance Rating, offers the system design engineer guidance as to how the fan was tested and rated. AMCA Publication 201 Fans and Systems, helps provide guidance as to what effect the system and its connections to the fan have on fan performance.

2. Symbols and Subscripts

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.

Fans are tested in setups that simulate installations. The four standard installation types are as shown in Figure 3.1.

For symbols and subscripted symbols, see Table 2.1. For subscripts, see Table 2.2.

3. Fan Testing

1.1 Purpose

AMCA INSTALLATION TYPE A: Free Inlet, Free Outlet

This part of the AMCA Fan Application Manual includes general information about how fans are tested in the laboratory, and how their performance ratings are calculated and published. It also reviews some of the more important reasons for the "loss" of fan performance that may occur when the fan is installed in an actual system.

AMCA INSTALLATION TYPE B: Free Inlet, Ducted Outlet

Allowances, called System Effect Factors (SEF), are also given in this part of the manual. SEF must be taken into account by the system design engineer if a reasonable estimate of fan/system performance is to be determined.

1.2 Some limitations It must be appreciated that the System Effect Factors given in this manual are intended as guidelines and are, in general, approximations. Some have been obtained from research studies, others have been published previously by individual fan manufacturers, and many represent the consensus of engineers with considerable experience in the application of fans. Fans of different types and even fans of the same type, but supplied by different manufacturers, will not necessarily react with the system in exactly the same way. It will be necessary, therefore, to apply judgment based on actual experience in applying the SEF. The SEF represented in this manual assume that the fan application is generally consistent with the method of testing and rating by the manufacturer. Inappropriate application of the fan will result in SEF

AMCA INSTALLATION TYPE C: Ducted Inlet, Free Outlet

AMCA INSTALLATION TYPE D: Ducted Inlet, Ducted Outlet

Figure 3.1 - Standard Fan Installation Types

3.1 ANSI/AMCA Standard 210 Most fan manufacturers rate the performance of their products from tests made in accordance with ANSI/AMCA 210 Laboratory Methods of Testing Fans for Aerodynamic Performance Rating. The purpose 1

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AMCA INTERNATIONAL, INC.

Table 2.1 - Symbols and Subscripted Symbols SYMBOL

DESCRIPTION

UNITS OF MEASURE SI I-P

A

Area of cross section

m2

ft2

D

Diameter, impeller

mm

in.

D

Diameter, Duct

m

ft

H

Fan Power Input

kw

hp

H/T

Hub-to-Tip Ratio

Dimensionless

Kp

Compressibility Coefficient

Dimensionless

Cp

Loss Coefficient

Dimensionless

N

Speed of Rotation

rpm

rpm

Ps

Fan Static Pressure

Pa

in. wg

Pt

Fan Total Pressure

Pa

in. wg

Pv

Fan Velocity Pressure

Pa

in. wg

pb

Corrected Barometric Pressure

kPa

in. Hg

PL

Plane of Measurement

---

---

Q

Airflow

m3/s

ft3/min

Re

Fan Reynolds Number

SEF

System Effect Factor

Pa

in. wg

td

Dry-Bulb Temperature

°C

°F

tw

Wet-Bulb Temperature

°C

°F

μ

Air Viscosity

Pa•s

lbm/ft•s

V

Velocity

m/s

fpm

W

Power Input to Motor

watts

watts

ηs

Fan Static Efficiency

%

%

ηt

Fan Total Efficiency

%

%

ρ

Air Density

kg/m3

lbm/ft3

Dimensionless

Table 2.2 - Subscripts SUBSCRIPT

DESCRIPTION

a c x 1 2 3 5 6 8

Atmospheric conditions Converted Value Plane 0, 1, 2, ...as appropriate Fan Inlet Plane Fan Outlet Plane Pitot Traverse Plane Plane 5 (nozzle inlet station in chamber) Plane 6 (nozzle discharge station in chamber) Plane 8 (inlet chamber measurement station)

2

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AMCA 201-02 (R2011)

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 installation where uniform airflow conditions do not exist, the fan's performance will be measurably reduced. As illustrated in Figure 3.3 Plane 2, the velocity profile at the outlet of a fan is not uniform. The section of straight duct attached to the fan outlet controls the diffusion of the outlet airflow and establishes a more uniform velocity as shown in Figure 3.3 Plane X. The energy loss when a gas, such as air, passes through a sudden enlargement is related to the square of the velocity. Thus the ducted outlet with its more uniform velocity significantly reduces the loss at the point of discharge to the atmosphere. A manufacturer may test a fan with or without an inlet duct or outlet duct. For products licensed to use the AMCA Certified Ratings Seal, catalog ratings will state whether ducts were used during the rating tests. If the fans are not to be applied with the same duct(s) as in the test setup, an allowance should be made for the difference in performance that may result.

1

Transition Piece

2

Straightener

FOR FAN INSTALLATION TYPES: B: Free Inlet, Ducted Outlet D: Ducted Inlet, Ducted Outlet Figure 3.2 - Pitot Traverse in Outlet Duct 3

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AMCA 201-02 (R2011)

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.

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)

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.

Psc = Ptc - Pvc

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.

ηtc = (Qc × Ptc × Kp) / (6362 • Hc)

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 4

Hc = H × (Dc/D)5 × (Nc/N)3 × (ρc/ρ) × (Kp/Kpc)

ηtc = (Qc × Ptc × Kp) / Hc

(SI) (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

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AMCA 201-02 (R2011)

BLAST AREA

PL X

PL 2

DISCHARGE DUCT OUTLET AREA

CUTOFF

CENTRIFUGAL FAN

PL 2

PL X

AXIAL FAN Figure 3.3 - Controlled Diffusion and Establishment of a Uniform Velocity Profile in a Straight Length of Outlet Duct

38mm ±6mm (1.5in. ±0.25 in.) PL.5 PL.6

PL.8

PL.1

PL.2 0.5 M MIN.

0.2M MIN.

0.5M MIN.

0.2 M MIN. 0.3 M MIN.

t d2 AIRFLOW

VARIABLE SUPPLY SYSTEM

M

FAN

t d3 SETTLING MEANS

0.1 M MIN.

SETTLING MEANS (See note 4)

Ps5

ΔP

Pt8

Figure 3.4 - Inlet Chamber Setup - Multiple Nozzles in Chamber (ANSI/AMCA 210-99, Figure 15) 5

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AMCA 201-02 (R2011)

ANSI/AMCA Standard 99-2404-03

Page 1 of 2

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

AMCA Drive Arrangement

1 SWSI

ISO 13349 Drive Arrangement

1 or 12 (Arr. 1 with sub-base)

Description

Fan Configuration

Alternative Fan Configuration

For belt or direct drive. Impeller overhung on shaft, two bearings mounted on pedestal base. Alternative: Bearings mounted on independant pedestals, with or without inlet box.

2 SWSI

For belt or direct drive.

2

Impeller overhung on shaft, bearings mounted in bracket supported by the fan casing. Alternative: With inlet box.

3 SWSI

3 or 11 (Arr. 3 with sub-base)

For belt or direct drive. Impeller mounted on shaft between bearings supported by the fan casing. Alternative: Bearings mounted on independent pedestals, with or without inlet box.

3 DWDI

6 or 18 (Arr. 6 with sub-base)

For belt or direct drive. Impeller mounted on shaft between bearings supported by the fan casing. Alternative: Bearings mounted on independent pedestals, with or without inlet boxes.

4 SWSI

4

For direct drive. Impeller overhung on motor shaft. No bearings on fan. Motor mounted on base. Alternative: With inlet box.

5 SWSI

5

For direct drive. Impeller overhung on motor shaft. No bearings on fan. Motor flange mounted to casing. Alternative: With inlet box.

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

Figure 3.5 - AMCA Standard 99-2404 / Page 1 6

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AMCA 201-02 (R2011)

AMCA 201-02 (R2011)

Page 2 of 2

ANSI/AMCA Standard 99-2404-03

AMCA Drive Arrangement

ISO 13349 Drive Arrangement

7 SWSI

7

Description

Fan Configuration

Alternative Fan Configuration

For coupling drive. Generally the same as Arr. 3, with base for the prime mover. Alternative: Bearings mounted on independent pedestals with or without inlet box.

7DWDI

17 (Arr. 6 with base for motor)

For coupling drive. Generally the same as Arr. 3 with base for the prime mover. Alternative: Bearings mounted on independent pedestals with or without inlet box.

8 SWSI

8

For direct drive. Generally the same as Arr. 1 with base for the prime mover. Alternative: Bearings mounted on independent pedestals with or without inlet box.

9 SWSI

9

For belt drive. Impeller overhung on shaft, two bearings mounted on pedestal base. Motor mounted on the outside of the bearing base. Alternative: With inlet box.

10 SWSI

10

For belt drive. Generally the same as Arr. 9 with motor mounted inside of the bearing pedestal. Alternative: With inlet box.

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

Figure 3.6 - AMCA Standard 99-2404 / Page 2 7

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AMCA 201-02

ANSI/AMCA Standard 99-3404-03

Page 1 of 1

Drive Arrangements for Axial Fans An American National Standard - Approved by ANSI on June 10, 2003 Note: All fan orientations may be horizontal or vertical

AMCA Drive Arrangement 1

ISO 13349 Drive Arrangement 1 12 (Arr. 1 with sub-base)

Description

Alternative Fan Configuration

Fan Configuration

For belt or direct drive. Impeller overhung on shaft, two bearings mounted either upstream or downstream of the impeller. Alternative: Single stage or two stage fans can be supplied with inlet box and/or discharge evasé.

3

3 11 (Arr. 3 with sub-base)

For belt or direct drive. Impeller mounted on shaft between bearings on internal supports. Alternative: Fan can be supplied with inlet box, and/or discharge evasé.

4

4

For direct drive. Impeller overhung on motor shaft. No bearings on fan. Motor mounted on base or integrally mounted.

M

M

M

M

Alternative: With inlet box and/or with discharge evasé.

7

7

For direct drive. Generally the same as Arr. 3 with base for the prime mover.

M

M

Alternative: With inlet box and/or discharge evasé.

8

8

For direct drive. Generally the same as Arr. 1 with base for the prime mover.

M

M

Alternative: Single stage or two stage fans can be supplied with inlet box and/or discharge evasé.

9

9

For belt drive. Generally same as Arr. 1 with motor mounted on fan casing, and/or an integral base. Alternative: With inlet box and/or discharge evasé

M

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

Figure 3.7 - AMCA Standard 99-3404 / Page 1 8

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AMCA 201-02 (R2011)

has adopted the term Fan Reynolds Number. Re = (πND2ρ) / (60μ) where: N = impeller rotational speed, rpm D = impeller diameter, m(ft) ρ = air density, kg/m3 (lbm/ft3) μ = absolute viscosity, 1.8185 × 10-5 Pa•s (5°C to 38°C) (SI) (1.22 × 10-5 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 Fan Total Pressure Fan Power Fan Static Efficiency Fan Total Efficiency

Ps Pt H ηs ηt

Air density (ρ), fan size (D), and fan rotational speed (N) are usually constant for the entire curve and must be stated. A typical fan performance curve is shown in Figure 4.1. Figure 4.2 illustrates examples of performance curves for a variety of fan types.

9

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AMCA 201-02 (R2011)

10

ηt

40

H

AIRFLOW, Q

Figure 4.1 - Fan Performance Curve at N RPM

10 OPERATION AT STANDARD DENSITY

50

ηs

60

30

20

0

EFFICIENCY, η PERCENT

100

Ps

70

POWER, H

80

PRESSURE, P

SIZE 30 FAN AT N RPM

Pt

90

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AMCA 201-02 (R2011)

TYPE

BACKWARDINCLINED BACKWARDCURVED

HOUSING DESIGN

• Highest efficiency of all centrifugal fan designs. • Ten to 16 blades of airfoil contour curved away from direction of rotation. Deep blades allow for efficient expansion within blade passages • Air leaves impeller at velocity less than tip speed. • For given duty, has highest speed of centrifugal fan designs

• Scroll-type design for efficient conversion of velocity pressure to static pressure. • Maximum efficiency requires close clearance and alignment between wheel and inlet

• Efficiency only slightly less than airfoil fan. • Ten to 16 single-thickness blades curved or inclined away from direction of rotation • Efficient for same reasons as airfoil fan.

• Uses same housing configuration as airfoil design.

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

RADIAL

CENTRIFUGAL FANS

AIRFOIL

IMPELLER DESIGN

R

• Scroll. Usually narrowest of all centrifugal designs. • Because wheel design is less efficient, housing dimensions are not as critical as for airfoil and backward-inclined fans.

M

CENTRIFUGAL CENTRIFUGAL

A

B

AXIAL

POWER ROOF VENTILATORS

SPECIAL DESIGNS

TUBULAR

VANEAXIAL

TUBEAXIAL

AXIAL FANS

PROPELLER

FORWARDCURVED

M

• Flatter pressure curve and lower efficiency than the airfoil, backward-curved, and backward-inclined. • Do not rate fan in the pressure curve dip to the left of peak pressure. • Power rises continually toward free delivery. Motor selection must take this into account.

• Scroll similar to and often identical to other centrifugal fan designs. • Fit between wheel and inlet not as critical as for airfoil and backward-inclined fans.

• Low efficiency. • Limited to low-pressure applications. • Usually low cost impellers have two or more blades of single thickness attached to relatively small hub. • Primary energy transfer by velocity pressure.

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

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

• Cylindrical tube with close clearance to blade tips.

• Good blade design gives medium- to high-pressure capability at good efficiency. • Most efficient of these fans have airfoil blades. • Blades may have fixed, adjustable, or controllable pitch. • Hub is usually greater than half fan tip diameter.

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

• Performance similar to backward-curved fan except capacity and pressure are lower. • Lower efficiency than backward-curved fan. • Performance curve may have a dip to the left of peak pressure.

• Cylindrical tube similar to vaneaxial fan, except clearance to wheel is not as close. • Air discharges radially from wheel and turns 90° to flow through guide vanes.

• Low-pressure exhaust systems such as general factory, kitchen, warehouse, and some commercial installations. • Provides positive exhaust ventilation, which is an advantage over gravity-type exhaust units. • Centrifugal units are slightly quieter than axial units.

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

• Low-pressure exhaust systems such as general factory, kitchen, warehouse, and some commercial installations. • Provides positive exhaust ventilation, which is an advantage over gravity-type exhaust units.

• Essentially a propeller fan mounted in a supporting structure • Hood protects fan from weather and acts as safety guard. • Air discharges from annular space at bottom of weather hood.

Figure 4.2 - Types of Fans Adapted with permission from 1996 ASHRAE Systems and Equipment Handbook (SI) 11

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AMCA 201-02 (R2011)

PERFORMANCE CURVES

a

PERFORMANCE CHARACTERISTICS

Pt

8

Ps

10

6 ηt 4

8 6

ηs wo

2

4 2

VOLUME FLOW RATE, Q

0 0

2

EFFICIENCY

PRESSURE-POWER

10

4

6

8

8 6

4

4 2

2

VOLUME FLOW RATE, Q 2

4

6

8

10

6

8 6

4

4 2

2

VOLUME FLOW RATE, Q 0

2

4

6

8

EFFICIENCY

PRESSURE-POWER

8

0

10

6

8 6

4

4 2

2

VOLUME FLOW RATE, Q 0

2

4

6

8

EFFICIENCY

PRESSURE-POWER

8

0

0 10

8

10

6

8 6

4

4 2

2

VOLUME FLOW RATE, Q

0 0

2

4

6

8

EFFICIENCY

PRESSURE-POWER

10

• Pressure curve less steep than that of backward-curved fans. Curve dips to left of peak pressure. • Highest efficiency to right of peak pressure at 40 to 50% of wide open volume. • Rate fan to right of peak pressure. • Account for power curve, which rises continually toward free delivery, when selecting motor.

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

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

• For low-pressure, high-volume air moving applications, such as air circulation in a space or ventilation through a wall without ductwork. • Used for makeup air applications.

• High flow rate, medium-pressure capabilities. • Performance curve dips to left of peak pressure. Avoid operating fan in this region. • Discharge pattern circular and airstream rotates or swirls.

• Low- and medium-pressure ducted HVAC applications where air distribution downstream is not critical. • Used in some industrial applications, such as drying ovens, paint spray booths, and fume exhausts.

• High-pressure characteristics with medium-volume flow capabilities. • Performance curve dips to left of peak pressure due to aerodynamic stall. Avoid operating fan in this region. • Guide vanes correct circular motion imprated by wheel and improve pressure characteristics and efficiency of fan.

• General HVAC systems in low-, medium-, and high-pressure applications where straight-through flow and compact installation are required. • Has good downstream air distribution • Used in industrial applications in place of tubeaxial fans. • More compact than centrifugal fans for same duty.

• Performance similar to backward-curved fan, except capacity and pressure is lower. • Lower efficiency than backward-curved fan because air turns 90°. • Performance curve of some designs is similar to axial flow fan and dips to left of peak pressure.

• Primarily for low-pressure, return air systems in HVAC applications. • Has straight-through flow.

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

• Low-pressure exhaust systems, such as general factory, kitchen, warehouse, and some commercial installations. • Low first cost and low operating cost give an advantage over gravity flow exhaust systems. • Centrifugal units are somewhat quieter than axial flow units.

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

• Low-pressure exhaust systems, such as general factory, kitchen, warehouse, and some commercial installations. • Low first cost and low operating cost give an advantage over gravity flow exhaust systems.

0 10

10 8

10

6

8 6

4

4 2

2

VOLUME FLOW RATE, Q

0 0

2

4

6

8

EFFICIENCY

PRESSURE-POWER

• Primarily for materials handling in industrial plants. Also for some high-pressure industrial requirements. • Rugged wheel is simple to repair in the field. Wheel sometimes coated with special material. • Not common for HVAC applications.

0 10

10

0 10

10 8

10

6

8 6

4

4 2

2

VOLUME FLOW RATE, Q

0 0

2

4

6

8

EFFICIENCY

PRESSURE-POWER

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

0 10

10

0 10

10 8

10

6

8 6

4

4 2

2

VOLUME FLOW RATE, Q

0 0

2

4

6

8

EFFICIENCY

PRESSURE-POWER

• Same heating, ventilating, and air-conditioning applications as airfoil fan. • Used in some industrial applications where airfoil blade may corrode or erode due to environment.

EFFICIENCY

PRESSURE-POWER

6

0

0 10

10 8

10

6

8 6

4

4 2

2 VOLUME FLOW RATE, Q

0 0

2

4

6

8

EFFICIENCY

PRESSURE-POWER

• Similar to airfoil fan, except peak efficiency slightly lower. 10

0

0 10

10 8

10

6

8 6

4

4 2

2

0 0

VOLUME FLOW RATE, Q 2 4 6

8

EFFICIENCY

PRESSURE-POWER

• General heating, ventilating, and air-conditioning applications. • Usually only applied to large systems, which may be low-, medium-, or high-pressure applications. • Applied to large, clean-air industrial operations for significant energy savings.

0 10

10 8

APPLICATIONS

• Highest efficiencies occur at 50 to 60% of wide open volume. This volume also has good pressure characteristics. • Power reaches maximum near peak efficiency and becomes lower, or self-limiting, toward free delivery.

0 10

a: These performance curves reflect general characteristics of various fans as commonly applied. They are not intended to provide complete selection criteria, since other parameters, such as diameter and speed, are not defined.

Figure 4.2 - Types of Fans Adapted with permission from 1996 ASHRAE Systems and Equipment Handbook (SI) 12

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AMCA 201-02 (R2011)

5.1 Type A: Free inlet, free outlet fans

1) Type B: Free inlet, ducted outlet 2) Type C: Ducted inlet, free outlet 3) Type D: Ducted inlet, ducted outlet

Fans designed for use other than with duct systems are usually rated over a lower range of pressures. They are commonly cataloged and sold as a complete unit with suitable drive and motor.

The performance of fans intended for use with duct systems is usually published in the form of a "multirating" table. A typical multi-rating table, as illustrated in Figure 5.2 shows:

Typical fans in this group are propeller fans and power roof ventilators. They are usually available in direct or belt-drive arrangements and performance ratings are published in a modified form of the multirating table. Figure 5.1 illustrates such a table for part of a line of belt-drive propeller fans.

a) the speed (N) in rpm b) the power (H) in kw (hp) c) the fan static pressure (Ps) in Pa (in. wg) d) the outlet velocity (V) in m/s, (fpm) e) the airflow (Q) in m3/s (cfm)

5. Catalog Performance Tables

5.2 Ducted fans There are three types of ducted fans, as described in Section 3: SIZE No. of Motor Peak rpm (cm) Blades kW kW 0.19 862 0.13 0.19 960 0.20 61 3 0.25 1071 0.27 0.37 1220 0.40 0.19 806 0.20 0.25 883 0.27 69 3 0.37 1035 0.43 0.56 1165 0.62 0.37 825 0.42 0.56 945 0.62 0.75 1045 0.82 84 3 1.12 1190 1.19 1.49 1306 1.64 TYPICAL RATING TABLE

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.

AIRFLOW (m3/s) @ STATIC PRESSURE (Pa) 0 31 62 93 124 155 186 217 2.02 1.58 0.58 2.25 1.87 0.97 2.51 2.18 1.76 0.76 2.86 2.57 2.24 1.70 0.81 2.89 2.36 1.05 3.17 2.68 1.94 0.76 3.71 3.30 2.85 1.56 0.95 4.18 3.83 3.44 3.01 1.60 1.10 4.36 3.76 3.04 1.27 4.99 4.48 3.92 2.38 1.42 5.23 5.08 4.57 4.01 2.31 1.52 6.29 5.90 5.47 5.01 4.48 2.79 1.94 6.91 6.53 6.15 5.75 5.32 4.81 3.05 2.24 FOR A SERIES OF BELT-DRIVEN PROPELLER FANS

248

1.84

SIZE No. of Motor Peak AIRFLOW (ft3/min) @ STATIC PRESSURE (in. wg) rpm (in.) Blades hp bhp 0 1/8 1/4 3/8 1/2 5/8 3/4 7/8 1 1/4 862 0.18 4,283 3,350 1,230 1/4 960 0.27 4,770 3,960 2,050 24 3 1/3 1071 0.36 5,321 4,620 3,730 1,600 1/2 1220 0.54 6,062 5,450 4,750 3,600 1,710 1/4 806 0.27 6,123 4,990 2,230 1/3 883 0.36 6,708 5,675 4,100 1,620 27 3 1/2 1035 0.57 7,862 7,000 6,035 3,315 2,020 3/4 1165 0.83 8,850 8,110 7,290 6,385 3,400 2,330 1/2 825 0.56 9,240 7,970 6,430 2,700 3/4 945 0.83 10,580 9,500 8,300 5,040 3,010 1 1045 1.1 11,710 10,755 9,685 8,490 4,890 3,215 33 3 1½ 1190 1.6 13,335 12,490 11,580 10,610 9,500 5,905 4,100 2 1306 2.2 14,630 13,845 13,030 12,185 11,280 10,200 6,470 4,740 3,900 TYPICAL RATING TABLE FOR A SERIES OF BELT-DRIVEN PROPELLER FANS Figure 5.1 - Propeller Fan Performance Table 13

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AMCA 201-02 (R2011)

IMPELLER DIAMETER: TIP SPEED IN m/s: Volume m3/s 1.81 2.17 2.53 2.89 3.25 3.61 3.97 4.33 4.69 5.06 5.42 5.78 6.14 6.50 6.86 7.22 7.94 8.67 9.39 10.11 10.83 11.55 12.28 13.00 13.72 14.44

Outlet Vel. (m/s) 2.55 3.06 3.56 4.07 4.58 5.08 5.59 6.10 6.61 7.13 7.63 8.14 8.65 9.15 9.66 10.17 11.18 12.21 13.23 14.24 15.25 16.27 17.30 18.31 19.32 20.34

62 Pa rpm

kW

222 236 253 272 292 314 338 361 385 409 434 458 483 508

0.14 0.17 0.22 0.27 0.34 0.42 0.51 0.62 0.74 0.88 1.03 1.21 1.41 1.63

927 mm .0485 × RPM 93 Pa rpm

kW

270 284 300 317 337 358 379 402 426 449 473 498 522 547 571 621

0.25 0.30 0.36 0.43 0.52 0.62 0.74 0.87 1.01 1.18 1.37 1.58 1.81 2.06 2.34 2.99

OUTLET AREA: MAXIMUM kW:

124 Pa rpm

kW

313 327 343 360 378 398 419 441 464 488 511 535 559 585 633 682

0.39 0.45 0.53 0.63 0.73 0.86 1.00 1.16 1.33 1.53 1.75 1.99 2.25 2.54 3.20 3.98

155 Pa rpm

kW

352 366 382 399 417 437 457 479 501 525 538 571 595 644 693 742 791

0.55 0.64 0.74 0.85 0.98 1.13 1.30 1.49 1.69 1.92 2.16 2.44 2.74 3.41 4.20 5.13 6.20

.71 SQ METERS 13.65 × (RPM/1000)3

186 Pa

217 Pa

rpm

kW

rpm

kW

389 403 419 436 454 473 494 515 537 560 584 607 654 703 752 801 850

0.75 0.86 0.98 1.11 1.26 1.44 1.63 1.86 2.09 2.35 2.62 2.93 3.63 4.44 5.38 6.47 7.70

411 424 438 455 472 489 509 529 550 572 595 616 665 712 761 810 859 908

0.87 0.98 1.10 1.25 1.41 1.58 1.79 2.01 2.26 2.54 2.82 3.14 3.85 4.68 5.64 6.73 7.99 9.40

246 Pa rpm

kW

443 458 472 489 506 524 543 564 585 606 629 675 721 769 818 867 916 965 1015

1.10 1.19 1.39 1.56 1.74 1.95 2.18 2.43 2.71 3.01 3.34 4.07 4.93 5.90 7.01 8.27 9.70 11.30 13.06

310 Pa

373 Pa

rpm

kW

rpm

kW

494 507 522 538 555 572 590 610 630 651 695 741 788 834 883 932 981 1030 1072 1129

1.52 1.68 1.86 2.06 2.28 2.53 2.78 3.07 3.39 3.74 4.52 5.40 6.41 7.57 8.87 10.32 11.95 13.77 15.78 17.98

540 554 568 584 600 617 635 654 674 715 759 805 852 898 946 995 1044 1093 1142

1.99 2.18 2.39 2.62 2.89 3.16 3.45 3.78 4.15 4.96 5.89 6.94 8.11 9.47 10.96 12.62 14.46 16.50 18.76

TYPICAL MULTISPEED RATING TABLE FOR A SINGLE WIDTH, SINGLE INLET CENTRIFUGAL FAN IMPELLER DIAMETER: TIP SPEED IN FPM: Volume CFM 3825 4590 5355 6120 6885 7650 8415 9180 9945 10710 11475 12240 13005 13770 14535 15300 16830 18360 19890 21420 22950 24480 26010 27540 29070 30600

Outlet Vel. (fpm) 500 600 700 800 900 1000 1100 1200 1300 1400 1500 1600 1700 1800 1900 2000 2200 2400 2600 2800 3000 3200 3400 3600 3800 4000

36.5 IN 9.56 × RPM

OUTLET AREA: MAXIMUM BHP:

7.65 SQ FT 18.3 × (RPM/1000)3

1/4 in. wg

3/8 in. wg

1/2 in. wg

5/8 in. wg

3/4 in. wg

7/8 in. wg

rpm

bhp

rpm

bhp

rpm

bhp

rpm

bhp

rpm

rpm

222 236 253 272 292 314 338 361 385 409 434 458 483 508

0.185 0.233 0.292 0.365 0.450 0.560 0.682 0.826 0.989 1.175 1.387 1.626 1.895 2.191

270 284 300 317 337 358 379 402 425 449 473 498 522 547 571 621

0.334 0.400 0.483 0.579 0.695 0.832 0.988 1.163 1.360 1.587 1.837 2.115 2.424 2.767 3.144 4.003

313 327 343 360 378 398 419 441 464 488 511 535 559 585 633 682

0.519 0.608 0.716 0.840 0.981 1.149 1.340 1.553 1.780 2.048 2.346 2.665 3.017 3.403 4.289 5.335

352 366 383 399 417 437 457 479 501 525 538 571 595 644 693 742 791

0.743 0.856 0.992 1.144 1.314 1.514 1.741 1.993 2.269 2.570 2.901 3.275 3.672 4.577 5.632 6.885 8.308

bhp

389 1.01 403 1.15 419 1.31 436 1.49 454 1.69 473 1.93 494 2.19 515 2.49 537 2.80 560 3.15 584 3.52 607 3.93 654 4.87 703 5.96 752 7.22 801 8.67 850 10.32

bhp

1 in. wg rpm

411 1.17 424 1.31 443 438 1.48 458 455 1.68 472 472 1.89 489 489 2.12 506 509 2.40 524 529 2.70 543 550 3.03 564 572 3.40 585 595 3.78 606 618 4.21 629 665 5.16 675 712 6.28 721 761 7.56 769 810 9.03 818 859 10.71 867 908 12.50 916 965 1015

bhp

1.48 1.60 1.86 2.09 2.34 2.61 2.92 3.26 3.64 4.04 4.48 5.46 6.61 7.91 9.40 11.09 13.01 15.16 17.52

1-1/4 in. wg 1-1/2 in. wg rpm

bhp

rpm

bhp

494 507 522 538 555 572 590 610 630 651 695 741 788 834 883 932 981 1030 1072 1129

2.04 2.25 2.49 2.76 3.06 3.39 3.73 4.12 4.55 5.02 6.06 7.24 8.60 10.15 11.89 13.84 16.03 18.47 21.16 24.11

540 554 568 584 600 617 635 654 674 715 759 805 852 898 946 995 1044 1093 1142

2.67 2.92 3.20 3.52 3.87 4.24 4.63 5.07 5.56 6.65 7.90 9.30 10.88 12.70 14.70 16.92 19.39 22.13 25.16

TYPICAL MULTISPEED RATING TABLE FOR A SINGLE WIDTH, SINGLE INLET CENTRIFUGAL FAN Figure 5.2 - Centrifugal Fan Performance Tables 14

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AMCA 201-02 (R2011)

OUTLET VELOCITY

500 600 700 800 900

1000 1100 1200 1300 1400

1500 1600 1700 1800 1900

2000 2200 2400 2600 2800

3000 3200 3400 3600 3800 4000

VOLUME CFM

3825 4590 5355 6120 6885

7650 8415 9180 9945 10710

11475 12240 13005 13770 14535

15300 16830 18360 19890 21420

22950 24480 26010 27540 29070 30600

.185 .233 .292 .365 .450

BHP

434 1.387 456 1.626 482 508 2.19

.334 .400 .483 .579

BHP

571 3.744 629 4.003

449 1.587 473 1.837 493 2.115 522 2.424 547 2.767

BHP

584 3.403 633 4.289 682 5.335

464 1.78 488 2.048 511 2.346 535 2.665 559 3.017

360 .840 378 .981 398 1.149 419 1.340 441 1.553

313 .519 327 .608 343 .716

RPM

1/2” SP

BHP

596 644 4.577 693 5.632 742 6.885 791 8.308

479 1.995 501 2.269 525 2.570 538 2.901 571 3.276

332 .992 399 1.144 417 1.314 437 1.514 457 1.741

352 .743 366 .856

RPM

5/8” SP

RPM

3.93 4.87 5.76 7.22 8.67

2.19 2.49 2.80 3.15 3.52

1.15 1.31 1.49 1.69 1.93

850 10.32

607 654 703 752 801

494 515 537 560 584

403 419 436 454 473

389 1.01

BHP

3/4” SP

390 RPM

337 .695 358 .822 379 .988 482 1.163 426 1.360

270 284 300 317

RPM

3/8” SP

BHP

4.21 5.16 6.28 7.56 9.03

2.40 2.70 3.03 3.40

1.31 1.48 1.58 1.89 2.12

859 10.71 908 12.60

618 665 712 761 810

509 529 550 572 595

424 438 455 472 489

411 1.17

RPM

7/8” SP

BHP

4.48 5.46 6.81 7.91 8.48

2.61 2.92 3.26 3.84 4.04

1.48 1.60 1.86 2.09 2.34

3.06 3.49 3.73 4.12 4.55

2.04 2.25 2.49 2.76

BHP

651 5.02 695 6.06 741 7.24 788 8.60 834 10.15

555 572 590 610 630

494 507 522 538

RPM

1-1/4” SP

BHP

3.52 3.87 4.24 4.63 5.07 674 5.56 715 6.65 759 7.90 9.30 852 10.88

584 600 617 635 654

540 2.67 554 2.92 568 3.28

RPM

1-1/2” SP

BHP

3.99 4.36 4.76 5.18 5.63 696 6.11 736 7.24 778 822 10.02 867 11.65

612 627 643 661 678

584 3.37 598 3.66

RPM

1-3/4” SP

867 11.09 883 11.89 898 12.70 914 13.48 916 13.01 932 13.84 946 14.70 960 15.56 965 15.16 981 16.03 995 16.92 1009 17.83 1015 17.52 1030 18.47 1044 19.39 1057 20.35 1079 21.16 1093 22.13 1106 23.12 1129 24.11 1142 25.16 1155 26.18

629 675 721 769 818

524 543 564 585 606

443 458 472 489 506

RPM

1” SP

490 RPM

314 .560 338 .682 361 .826 335 .988 409 1.175

222 236 253 272 292

RPM

1/4” SP

PRESSURE IN IN. WG BRAKE HORSEPOWER

585 RPM 810 RPM

RECOMMENDED SELECTION RANGE

CFM

Figure 5.3 - Typical Fan Performance Table Showing Relationship to a Family of Constant Speed Performance Curves

15

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AMCA 201-02 (R2011)

Most performance tables do not cover the complete range from no delivery to free delivery but cover only the typical operating range. Figure 5.4 illustrates the recommended performance range of a centrifugal fan. Comparison of Figure 5.4 with Figure 5.3 will show that the published performance table also covers only the recommended performance range of the fan.

6. Air Systems 6.1 The system An air system may consist simply of a fan with ducting connected to either the inlet or outlet or to both. A more complicated system may include a fan, ductwork, air control dampers, cooling coils, heating coils, filters, diffusers, sound attenuation, turning vanes, etc. See AMCA Publication 200 Air Systems, for more information.

It should be remembered that fans are generally tested without obstructions in the inlet and outlet and without any optional airstream accessories in place. Catalog ratings will, therefore, usually apply only to the bare fan with unobstructed inlet and outlet.

6.2 Component losses

Fan performance adjustment factors for airstream accessories are normally available from either the fan catalog or the fan manufacturer.

Every system has a combined resistance to airflow that is usually different from every other system and is dependent upon the individual components in the system.

Fans are usually tested in arrangement 1, or similar (see Figure 3.5). Rating tables will, therefore, also apply only to the tested arrangement. Allowances for the effect of bearing supports used in other arrangements should be obtained from the manufacturer if not shown in the catalog.

The determination of the "pressure loss" or "resistance to airflow," for the individual components can be obtained from the component manufacturers. The determination of pressure losses for ductwork design is well documented in standard handbooks such as the ASHRAE Handbook of Fundamentals.

SELECTION NOT USUALLY RECOMMENDED IN THIS RANGE

RECOMMENDED SELECTION RANGE ES PR

SY

ST

EM

CU RV E

PRESSURE

RE SU

DU CT

E RV

EM

SELECTION NOT USUALLY RECOMMENDED IN THIS RANGE

CU

ST

Y TS

C

DU

AIRFLOW Figure 5.4 - Recommended Performance Range of a Typical Centrifugal Fan 16

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AMCA 201-02 (R2011)

In a later section, the effects of some system components and fan accessories on fan performance are discussed. The System Effects presented will assist the system designer to determine fan selection.

The system curve of a "fixed system" plots as a parabola in accordance with the above relationship. Typical plots of the resistance to flow versus volume airflow for three different and arbitrary fixed systems, (A, B, and C) are illustrated in Figure 6.1. For a fixed system an increase or decrease in airflow results in an increase or decrease in the system resistance along the given system curve only. Also, as the components in a system change, the system curve changes.

6.3 The system curve At a fixed airflow through a given air system a corresponding pressure loss, or resistance to this airflow, will exist. If the airflow is changed, the resulting pressure loss, or resistance to airflow, will also change. The relationship between airflow pressure and loss can vary as a function of type of duct components, their interaction and the local velocity magnitude. In many cases, typical duct systems operate in the turbulent flow regime and the pressure loss can be approximated as a function of velocity (or airflow) squared. The simplifying relationship used in this publication governing the change in pressure loss as a function of airflow for a fixed system is:

Refer to Figure 6.1, Duct System A. With a system at the design airflow (Q) and at a design system resistance (P), an increase in airflow to 120% of Q will result in an increase in system resistance P of 144% since system resistance varies with the square of the airflow. Likewise, a decrease in airflow Q to 50% would result in a decrease in system resistance P to 25% of the design system resistance. In Figure 6.1, System Curve B is representative of a system that has more component pressure loss than System Curve A, and System Curve C has less component pressure loss than System Curve A.

Pc/P = (Qc/Q)2

Notice that on a percentage basis, the same relationships also hold for System Curves B and C. These relationships are characteristic of typical fixed systems.

A more through discussion of duct system pressure losses can be found in AMCA Publication 200 Air Systems.

200

160 M

140

C

E ST

SY

120 100 EM

B

80

EM

ST

A

SYSTEM DESIGN POINT

SY

ST

60

SY

PERCENT OF SYSTEM RESISTANCE

180

40 20 0 0

20

40

60

80

100

120

140

160

180

200

PERCENT OF SYSTEM AIRFLOW Figure 6.1 - System Curves 17

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AMCA 201-02 (R2011)

is now at Point 3 (the intersection of the fan curve and the new System C), with the airflow at approximately 120% of Q.

6.4 Interaction of system curve and fan performance curve If the system characteristic curve, composed of the resistance to system airflow and the appropriate SEF have been accurately determined, then the fan will deliver the designated airflow when installed in the system.

6.5 Effect of changes in speed Increases or decreases in fan rotational speed will alter the airflow through a system. According to the Fan Laws (see below), the % increase in airflow is directly proportional to the fan rotational speed ratio, and the fan static pressure is proportional to the square of the fan rotational speed ratio. Thus, a 10% increase in fan rotational speed will result in a new fan curve with a 10% increase in Q, as illustrated in Figure 6.3. Since the system components did not change, System Curve A remains the same. With airflow increasing by 10% over the original Q, the system resistance increases along System Curve A to Point 2, at the intersection with the new fan curve.

The point of intersection of the system curve and the fan performance curve determines the actual airflow. System Curve A in Figure 6.2 has been plotted with a fan performance curve that intersects the system design point. The airflow through the system in a given installation may be varied by changing the system resistance. This is usually accomplished by using fan dampers, duct dampers, mixing boxes, terminal units, etc.

The greater airflow moved by the fan against the resulting higher system resistance to airflow is a measure of the increased work done. In the same system, the fan efficiency remains the same at all points on the same system curve. This is due to the fact that airflow, system resistance, and required power are varied by the appropriate ratio of the fan rotational speed.

200 EM

A

180

M

ST

160 140

C

E ST

SY

PERCENT OF SYSTEM RESISTANCE

Figure 6.2 shows the airflow may be reduced from design Q by increasing the resistance to airflow, i.e., changing the system curve from System A to System B. The new operating point is now at Point 2 (the intersection of the fan curve and the new System B) with the airflow at approximately 80% of Q. Similarly, the airflow can be increased by decreasing the resistance to airflow, i.e., changing the system curve from System A to System C. The new operating point

SY

FAN CURVE

120

2

100

1

80

SYSTEM DESIGN POINT

3

60

EM

B

T YS

40

S

20 0 0

20

40

60

80

100

120

140

160

PERCENT OF SYSTEM AIRFLOW

Figure 6.2 - Interaction of System Curves and Fan Curve

18

180

200

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AMCA 201-02 (R2011)

air density of 1.2 kg/m3 (0.075 lbm/ft3) is standard in the fan industry throughout the world. Figure 6.4 illustrates the effect on the fan performance of a density variation from the standard value.

6.5.1 Fan Laws - effect of change in speed - (fan size and air density remaining constant) For the same size fan, Dc = D and, therefore, (Dc/D) = 1. When the air density does not vary, ρc = ρ and the air density ratio (ρc/ρ) = 1. Kp is taken as equal to unity in this and following examples.

6.6.1 Fan Laws - effect of change in density - (fan size and speed remaining constant) When the speed of the fan does not change, Nc = N and, therefore, (Nc/N) = 1. The fan size is also fixed, Dc = D and therefore (Dc/D) = 1.

Qc = Q × (Nc/N) Ptc = Pt × (Nc/N)2 Psc = Ps × (Nc/N)2

Qc = Q

Pvc = Pv × (Nc/N)2

Ptc = Pt × (ρc/ρ) Psc = Ps × (ρc/ρ)

Hc = H × (Nc/N)3

Pvc = Pv × (ρc/ρ)

6.6 Effect of density on system resistance

Hc = H × (ρc/ρ)

SY CT

S (AT 1.1N)

PRESSURE

H (AT 1.1N) 133

140

S (AT N) PRESSURE

120

2 H (AT N)

100

1

100

80 60 50

40 20

PERCENT OF POWER

160

DU

PERCENT OF SYSTEM RESISTANCE

ST

EM

A

The resistance of a duct system is dependent upon the density of the air flowing through the system. An

110%

0 0

20

40

60

80

100

120

140

160

180

200

PERCENT OF SYSTEM AIRFLOW

Figure 6.3 - Effect of 10% increase in Fan Speed

19

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AMCA 201-02 (R2011)

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AMCA 201-02 (R2011)

PERCENT OF SYSTEM RESISTANCE AND FAN PRESSURE

FAN PRESSURE CURVE @ DENSITY ρ

SYSTEM A @ DENSITY ρ FAN INLET

100

SYSTEM A @ DENSITY ρ/2 FAN INLET

FAN PRESSURE CURVE @ DENSITY ρ/2

80 60 40 20 0

PERCENT OF POWER

100 POWER @ DENSITY ρ

80 60 40 POWER @ DENSITY ρ/2

20 0 0

20

40

60

80

100

120

140

PERCENT OF SYSTEM AIRFLOW

Figure 6.4 - Density Effect

20

160

180

200

6.7 Fan and system interaction When system pressure losses have been accurately estimated and desirable fan inlet and outlet conditions have been provided, design airflow can be expected, as illustrated in Figure 6.5. Note again that the intersection of the actual system curve and the fan curve determine the actual airflow. However, when system pressure losses have not been accurately estimated as in Figure 6.6, or when undesirable fan inlet and outlet conditions exist as in Figure 6.7, design performance may not be obtained.

6.8 Effects of errors in estimating system resistance 6.8.1 Higher system resistance. In Figure 6.6, System Curve B shows a situation where a system has greater resistance to airflow than designed (Curve A). This condition is generally a result of inaccurate allowances of system resistance. All pressure losses must be considered when calculating system resistance or the actual system will be more restrictive to airflow than intended. This

condition results in an actual airflow at Point 2, which is at a higher pressure and lower airflow than was expected. If the actual duct system pressure loss is greater than design, an increase in fan speed may be necessary to achieve Point 5, the design airflow. CAUTION: Before increasing fan rotational speed, check with the fan manufacturer to determine whether the fan rotational speed can be safely increased. Also determine the expected increase in power. Since the power required increases as the cube of the fan rotational speed ratio, it is very easy to exceed the capacity of the existing motor and that of the available electrical service. 6.8.2 Lower system resistance. Curve C in Figure 6.6 shows a system that has less resistance to airflow than designed. This condition results in an actual airflow at Point 3, which is at a lower pressure and higher airflow than was expected.

CALCULATED SYSTEM CURVE PEAK FAN PRESSURE

1

DESIGN RESISTANCE

FAN PRESSURE CURVE

DESIGN AIRFLOW

Figure 6.5 - Fan/System Curve at Design Point 21

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AMCA 201-02 (R2011)

the fan speed, adjusting the variable inlet vane (VIV), if installed, or inlet dampers. The system resistance could also be increased to Point 1 on Curve A, Figure 6.6. The change in fan operating point should be evaluated carefully, since a change in fan power consumption may occur.

6.9 Safety factors It has been common practice among system designers to add safety factors to the calculated system resistance to account for the “unexpected”. In some cases, safety factors may compensate for resistance losses that were unaccounted for and the actual system will deliver the design airflow, Point 1, Figure 6.6. If the actual system resistance is lower than the design system resistance, including the safety factors, the fan will run at Point 3 and deliver more airflow. This result may not be advantageous because the fan may be operating at a less efficient point on the fan’s performance curve and may require more power than a properly designed system. Under these conditions, it may be desirable to reduce the fan performance to operate at Point 4 on Curve C, Figure 6.6. This may be accomplished by reducing

The system designer should also evaluate the fan performance tolerance and system resistance tolerance to determine if the lower or upper limits of the probable airflow in the system are acceptable. The combination of these tolerances should be evaluated to ensure that the “high-side” system resistance curve does not fall into the unstable range of performance. Operation in this area of the curve should be avoided and precautions taken to ensure operations outside of the unstable area, especially at the highest expected system resistance.

CURVE B: ACTUAL SYSTEM

ACTUAL SYSTEM RESISTANCE MORE THAN DESIGN

CURVE A: CALCULATED SYSTEM

5 CURVE C ACTUAL SYSTEM PEAK FAN PRESSURE

2 1

DESIGN RESISTANCE

3

ACTUAL SYSTEM LESS THAN DESIGN

4

FAN PRESSURE CURVE

DESIGN AIRFLOW

Figure 6.6 - Fan/System Curve Not at Design Point 22

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AMCA 201-02 (R2011)

6.10 Deficient fan/system performance The most common causes of deficient fan/system performance are improper fan inlet duct design, fan outlet duct design, and fan installation into the duct system. Any one or a combination of these conditions that alter the aerodynamic characteristics of the air flowing through the fan such that the fan’s full airflow potential, as tested in the laboratory and cataloged, is not likely to be realized. Other major causes of deficient performance are: • The air performance characteristics of the installed system are significantly different from the system designer's intent (See Figure 6.6). This may be due to a change in the system by others or unexpected behavior of the system during operation. • The system design calculations did not include adequate allowances for the effect of accessories and appurtenances (See Section 10). • The fan selection was made without allowing for the effect of appurtenances on the fan's performance (See Section 10). • Dirty filters, dirty ducts, dirty coils, etc., will increase the system resistance, and consequently, reduce the airflow - often significantly. • The "performance" of the system has been determined by field measurement techniques that have a high degree of uncertainty. Other "on-site" problems are listed in AMCA Publication 202 Troubleshooting, which includes detailed checklists and recommendations for the correction of problems with the performance of air systems.

6.11 Precautions performance

to

prevent

deficient

• Use appropriate allowances in the design calculations when space or other factors dictate the use of less than optimum arrangement of the fan outlet and inlet connections (See Sections 8 and 9). • Design the connections between the fan and the system to provide, as nearly as possible, uniform airflow conditions at the fan outlet and inlet connections (See Sections 8 and 9).

• Include adequate allowance for the effect of all accessories and appurtenances on the performance of the system and the fan. If possible, obtain from the fan manufacturer data on the effect of installed appurtenances on the fan's performance (See Section 10). • Use field measurement techniques that can be applied effectively on the particular system. Be aware of the probable accuracy of measurement and conditions that affect this. Refer to AMCA Publication 203 Field Performance Measurement of Fan Systems; for more precise measurement see AMCA Standard 803 Industrial Process/Power Generation Fans: Site Performance Test Standard. Also, refer to AABC National Standards, Chapter 8, Volume Measurements, Associated Air Balance Council, 5th Edition, 1989.

6.12 System Effect Figure 6.7 illustrates deficient fan/system performance resulting from one or more of the undesirable airflow conditions listed in Section 6.10. It is assumed that the system pressure losses, shown in system curve A, have been accurately determined, and a suitable fan selected for operation at Point 1. However, no allowance has been made for the effect of the system connections on the fan's performance. To account for this System Effect it will be necessary to add a System Effect Factor (SEF) to the calculated system pressure losses to determine the actual system curve. The SEF for any given configuration is velocity dependent and will vary across a range of airflow. This will be discussed in more detail in Section 7. (See Figure 7.1). In Figure 6.7 the point of intersection between the fan performance curve and the actual system curve B is Point 4. The actual airflow will be deficient by the difference 1-4. To achieve design airflow, a SEF equal to the pressure difference between Point 1 and 2 should have been added to the calculated system pressure losses and the fan selected to operate at Point 2. Note that because the System Effect is velocity related, the difference represented between Points 1 and 2 is greater than the difference between Points 3 and 4. The System Effect includes only the effect of the system configuration on the fan's performance.

23

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7. System Effect Factor (SEF)

7.1 System Effect Curves

A System Effect Factor is a value that accounts for the effect of conditions adversely influencing fan performance when installed in the air system.

Figure 7.1 shows a series of 19 System Effect Curves. By entering the chart at the appropriate air velocity (on the abscissa), it is possible to read across from any curve (to the ordinate) to find the SEF for a particular configuration.

CURVE B ACTUAL SYSTEM WITH SYSTEM EFFECT

CURVE A CALCULATED SYSTEM WITH NO ALLOWANCE FOR SYSTEM EFFECT

2 SYSTEM EFFECT LOSS AT DESIGN AIRFLOW

4

DESIGN RESISTANCE

1 3

SYSTEM EFFECT AT ACTUAL AIRFLOW

FAN CATALOG PRESSURE CURVE

AIRFLOW DEFICIENCY

DESIGN AIRFLOW

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

24

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AMCA 201-02 (R2011)

300

T

200

U

V

W 100 90

80

70

SYSTEM EFFECT FACTOR PRESSURE, Pa

1000

30

20

9 10

8

7

6

5

4

3

20 2.5

FG H I J K L

M

N

O

P

Q R 500

S

60

X 50

40

30

AIR VELOCITY, (m/s)

(Air Density = 1.2 kg/m3)

Figure 7.1 - System Effect Curves (SI)

900

800

700

600

400

25

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AMCA 201-02 (R2011)

FG H I J K L

5.0

M

N

O P

4.0

Q 3.0 R

SYSTEM EFFECT FACTOR - PRESSURE, in. wg

2.5 2.0

S 1.5 T 1.0 0.9

U

0.8 0.7 0.6

V

0.5 0.4

W

0.3 0.25 X

0.2

0.15

0.1

5

6

7

8 9 10

15

20

25

30

AIR VELOCITY, ft/min × 100 (Air Density = 0.075 lbm/ft3)

Figure 7.1 - System Effect Curves (I-P)

26

40

50

60

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AMCA 201-02 (R2011)

Table 7.1 - System Effect Coefficients

Curve in Figure 7.1

Dynamic Pressure Loss Coefficient C

F G H I J K L M N O P Q R S T U V W X

16.00 14.20 12.70 11.40 9.50 7.90 6.40 4.50 3.20 2.50 1.90 1.50 1.20 0.75 0.50 0.40 0.25 0.17 0.10 2

⎛ V ⎞ SEF = C ⎜ ⎟ ρ ⎝ 1.414 ⎠

SI

2

⎛ V ⎞ SEF = C ⎜ ⎟ ρ ⎝ 1097 ⎠

I-P

27

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AMCA 201-02 (R2011)

The SEF is given in Pascals (in. wg) and must be added to the total system pressure losses as shown on Figure 7.2. The velocity used when entering Figure 7.1 will be either the inlet or the outlet velocity of the fan. This will depend on whether the configuration in question is related to the fan inlet or the fan outlet. Most catalog ratings include outlet velocity figures but, for centrifugal fans, it may be necessary to calculate the inlet velocity (See Figure 9.14). The inlet velocity and outlet velocity of an axial fan can be approximated by using the fan impeller diameter to determine the airflow area. The necessary dimensioned drawings are usually included in the fan catalog. In Sections 8 and 9, typical inlet and outlet configurations are illustrated and the appropriate System Effect Curve is listed for each configuration. If more than one configuration is included in a system, the SEF for each must be determined separately and the total of these System Effects must be added to the total pressure losses.

The System Effect Curves are plotted for standard air at a density of 1.2 kg/m3 (0.075 lbm/ft3). Since the System Effect is directly proportional to density, values for other densities can be calculated as below: ⎛d ⎞ SEF2 = SEF1 ⎜ 2 ⎟ ⎝ d1 ⎠ Where: SEF2 = SEF at actual density SEF1 = SEF at standard density d2 = actual density d1 = standard density Alternatively, the SEF may be calculated by the method shown in Table 7.1. Determine the configuration being evaluated and use the appropriate loss coefficient, Cp, and application velocity, V. The SEF can then be calculated using the equations shown in Table 7.1.

FAN POWER

ACTUAL SYSTEM RESISTANCE

ACTUAL POWER REQUIRED

ACTUAL SYSTEM W/ SEF

SEF

CALCULATED SYSTEM W/NO ALLOWANCE FOR SEF

FAN PRESSURE

DESIGN AIRFLOW

Figure 7.2 - Effect of System on Fan Selection 28

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AMCA 201-02 (R2011)

should examine catalog ratings carefully for statements defining whether the published ratings are based on tests made with A: free inlet, free outlet; B: free inlet, ducted outlet; C: ducted inlet, free outlet or D; ducted inlet, ducted outlet.

7.2 Power determination When all the applicable System Effect Factors (SEF) have been added to the calculated system pressure losses the power shown in the catalog for the actual point of operation, Figure 7.2 or Table 7.1 may be used without further adjustment.

8.1 Outlet ducts

ANSI/AMCA 210 specifies an outlet duct that is no greater than 105% or less than 95% of the fan outlet area. It also requires that the slope of the transition elements be no greater than 15° for converging elements or greater than 7° for diverging elements.

As previously discussed, fans intended primarily for use with duct systems are usually tested with an outlet duct in place (See Figure 3.2). In most cases it is not practical for the fan manufacturer to supply this duct as part of the fan, but rated performance will not be achieved unless a comparable duct is included in the system design. The system design engineer

Figure 8.1 shows changes in velocity profiles at various distances from centrifugal and axial flow fan outlets. By definition, 100% "effective duct length" is a minimum of two and one half (2½) equivalent duct diameters. For velocities greater than 13 m/s (2500 fpm), add 1 duct diameter for each additional 5 m/s (1000 fpm).

8. Outlet System Effect Factors

BLAST AREA DISCHARGE DUCT CUTOFF

OUTLET AREA

25% 50% CENTRIFUGAL FAN

75% 100% EFFECTIVE DUCT LENGTH

AXIAL FAN

To calculate 100% duct length, assume a minimum of 2½ duct diameters for 12.7 m/s (2500 fpm) or less. Add 1 duct diameter for each additional 5.08 m/s (1000 fpm). EXAMPLE: 25.4 m/s (5000 fpm) = 5 equivalent duct diameters. If the duct is rectangular with side dimensions a and b, the equivalent duct diameter is equal to (4ab/π)0.5. Figure 8.1 - Fan Outlet Velocity Profiles 29

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AMCA 201-02 (R2011)

8.1.1 Axial flow fan - outlet ducts. Most exhaust axial flow fans are tested and/or rated with two to three equivalent duct diameters attached to the fan outlet. Often, fans are installed without an outlet duct, either because of available space or for economic reasons. Tubeaxial fans installed with no outlet ducts have System Effect Factors (SEF) approaching zero. Vaneaxial fans, however, do not perform as cataloged when they are installed with less than 50% "effective duct length." System Effect Curves for tubeaxial and vaneaxial fans with less than optimum outlet duct are shown in Figure 8.2. To determine the applicable SEF, calculate the average velocity in the outlet duct and enter the System Effect Curve (Figure 7.1) at this velocity, utilizing the appropriate System Effect Curve selected from Figure 8.2, then read over horizontally to the System Effect Factor, Pascals (in. wg) on the ordinate. 8.1.2 Centrifugal flow fan - outlet ducts. Centrifugal fans are sometimes installed with a less than optimum outlet duct. If it is not possible to use a

full-length outlet duct, then a SEF must be added to the system resistance losses. System Effect Curves for centrifugal fans with less than optimum outlet duct length are shown in Figure 8.3.

8.2 Outlet diffusers Many air systems are space-constricted and must, of necessity, use relatively small ducts having high static pressure losses. If space is not severely constricted, the use of larger ductwork and moving air at a lower velocity may be beneficial. Larger ductwork (within reason) reduces system pressure requirements. To effectively transition from a smaller duct size to a larger duct size it is necessary to use a connection piece between the duct sections that allows the airstream to expand gradually. This piece is called a diffuser, or evasé. These terms are used interchangeably in the industry. A properly designed evasé has a smooth and gradual transition between the duct sizes so that airflow is relatively undisturbed. An evasé operates on a very simple principle: air flowing from the smaller area to the larger area loses

AXIAL FAN

100% EFFECTIVE DUCT LENGTH To calculate 100% duct length, assume a minimum of 2½ duct diameters for 12.7 m/s (2500 fpm) or less. Add 1 duct diameter for each additional 5.08 m/s (1000 fpm). EXAMPLE: 25.4 m/s (5000 fpm) = 5 equivalent duct diameters

No Duct

12% Effective Duct

25% Effective Duct

50 % Effective Duct

100% Effective Duct

Tubeaxial Fan

---

---

---

---

---

Vaneaxial Fan

U

V

W

---

---

Determine SEF by using Figure 7.1 Figure 8.2 - System Effect Curves for Outlet Ducts - Axial Fans 30

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AMCA 201-02 (R2011)

velocity as it approaches the larger area, and a portion of the change (reduction) in velocity pressure is converted into static pressure. This process is called “static regain”, and is simply defined as the conversion of velocity pressure to static pressure. The efficiency of conversion (or loss of total pressure) will depend upon the angle of expansion, the length of the evasé section, and the blast area/outlet area ratio of the fan. The fan manufacturer will, in most cases, be able to provide design information for an efficient diffuser.

See AMCA Publication 200 Air Systems, for an example showing the effect of a diffuser on a duct exit.

8.3 Outlet duct elbows Values for pressure losses through elbows, which are published in handbooks and textbooks, are based upon a uniform velocity profile at entry into the elbow. Any non-uniformity in the velocity profile ahead of the elbow will result in a pressure loss greater than the industry-accepted value.

BLAST AREA DISCHARGE DUCT OUTLET AREA

CUTOFF

100% EFFECTIVE DUCT LENGTH

CENTRIFUGAL FAN To calculate 100% duct length, assume a minimum of 2½ duct diameters for 2500 fpm or less. Add 1 duct diameter for each additional 1000 fpm. EXAMPLE: 5000 fpm = 5 equivalent duct diameters. If the duct is rectangular with side dimensions a and b, the equivalent duct diameter is equal to (4ab/π)0.5.

Pressure Recovery

No Duct

12% Effective Duct

25% Effective Duct

50% Effective Duct

100% Effective Duct

0%

50%

80%

90%

100%

W W W-X — — — —

— — — — — — —

Blast Area Outlet Area 0.4 0.5 0.6 0.7 0.8 0.9 1.0

System Effect Curve P P R-S S T-U V-W —

R-S R-S S-T U V-W W-X —

U U U-V W-X X — —

Determine SEF by using Figure 7.1 Figure 8.3 - System Effect Curves for Outlet Ducts - Centrifugal Fans 31

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AMCA 201-02 (R2011)

Since the velocity profile at the outlet of a fan is not uniform, an elbow located at or near the fan outlet will develop a pressure loss greater than the industryaccepted value.

8.3.1 Axial fans - outlet duct elbows. Tubeaxial fans with two-piece and four-piece mitered elbows at varying distances from the fan outlet have a negligible SEF (see Figure 8.4).

The amount of this loss will depend upon the location and orientation of the elbow relative to the fan outlet. In some cases, the effect of the elbow will be to further distort the outlet velocity profile of the fan. This will increase the losses and may result in such uneven airflow in the duct that branch- takeoffs near the elbow will not deliver their design airflow. (See Section 8.6)

Vaneaxial fans with two and four-piece mitered elbows at varying distances from the fan outlet resulted in System Effect Curves as shown in Figure 8.4.

Wherever possible, a length of straight duct should be installed at the fan outlet to permit the diffusion and development of a uniform airflow profile before an elbow is inserted in the duct. If an elbow must be located near the fan outlet then it should be a radius elbow having a minimum radius-to-duct-diameter ratio of 1.5.

8.3.2 Centrifugal fans - outlet duct elbows. The outlet velocity of centrifugal fans is generally higher toward one or adjacent sides of the rectangular duct. If an elbow must be located near the fan outlet it should have a minimum radius-to-duct-diameter ratio of 1.5, and it should be arranged to give the most uniform airflow possible. Figure 8.5 gives System Effect Curves that can be used to estimate the effect of an elbow at the fan outlet. It also shows the reduction in losses resulting from the use of a straight outlet duct.

TUBEAXIAL FAN SHOWN

% EFFECTIVE DUCT LENGTH

% EFFECTIVE DUCT LENGTH

VANEAXIAL FAN SHOWN

90° Elbow

No Duct

12% Effective Duct

25% Effective Duct

50 % Effective Duct

100% Effective Duct

Tubeaxial Fan

2 & 4 Pc

---

---

---

---

---

Vaneaxial Fan

2 Pc

U

U-V

V

W

---

Vaneaxial Fan

4 Pc

W

---

---

---

---

Determine SEF by using Figures 7.1 and 8.1 Figure 8.4 - System Effect Curves for Outlet Duct Elbows - Axial Fans 32

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AMCA 201-02 (R2011)

POSITION C

POSITION D

POSITION B

E TIV TH C G FE EF LEN % CT DU

INL

ET

POSITION A

SWSI CENTRIFUGAL FAN SHOWN

Note: Fan Inlet and elbow positions must be oriented as shown for the proper application of the table on the facing page. Figure 8.5 - Outlet Elbows on SWSI Centrifugal Fans

33

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AMCA 201-02 (R2011)

Outlet Elbow Position

No Outlet Duct

12% Effective Duct

25% Effective Duct

50% Effective Duct

0.4

A B C D

N M-N L-M L-M

O N M M

P-Q O-P N N

S R-S Q Q

0.5

A B C D

O-P N-O M-N M-N

P-Q O-P N N

R Q O-P O-P

T S-T R-S R-S

0.6

A B C D

Q P N-O N-O

Q-R Q O O

S R Q Q

U T S S

0.7

A B C D

R-S Q-R P P

S R-S Q Q

T S-T R-S R-S

V U-V T T

0.8

A B C D

S R-S Q-R Q-R

S-T S R R

T-U T S S

W V U-V U-V

0.9

A B C D

T S R R

T-U S-T S S

U-V T-U S-T S-T

W W V V

1.0

A B C D

T S-T R-S R-S

T-U T S S

U-V U T T

W W V V

SYSTEM EFFECT CURVES FOR SWSI FANS

DETERMINE SEF BY USING FIGURES 7.1 AND 8.1 For DWDI fans determine SEF using the curve for SWSI fans. Then, apply the appropriate multiplier from the tabulation below MULTIPLIERS FOR DWDI FANS ELBOW POSITION A = ∆P × 1.00 ELBOW POSITION B = ∆P × 1.25 ELBOW POSITION C = ∆P × 1.00 ELBOW POSITION D = ∆P × 0.85 Figure 8.5 - Outlet Elbows on SWSI Centrifugal Fans 34

100% Effective Duct

NO System Effect Factor

Blast Area Outlet Area

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AMCA 201-02 (R2011)

8.4 Turning vanes

a large plenum or to free space a parallel blade damper may be satisfactory.

Turning vanes will usually reduce the pressure loss through an elbow, however, where a non-uniform approach velocity profile exists, such as at a fan outlet, the vanes may serve to continue the nonuniform profile beyond the elbow. This may result in increased losses in other system components downstream of the elbow.

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.

8.5 Volume control dampers Volume control dampers are manufactured with either "opposed" blades or "parallel" blades. When partially closed, the parallel bladed damper diverts the airstream to the side of the duct. This results in a non-uniform velocity profile beyond the damper and airflow to branch ducts close to the downstream side may be seriously affected. The use of an opposed blade damper is recommended when air volume control is required at the fan outlet and there are other system components, such as coils or branch takeoffs downstream of the fan. When the fan discharges into

PARALLEL-BLADE DAMPER ILLUSTRATING DIVERTED AIRFLOW

When a damper is required, it is often furnished as accessory equipment by the fan manufacturer (see Figure 8.6). In many systems, a volume control damper will be located in the ductwork at or near the fan outlet. Published pressure drops for wide-open control dampers are based on uniform approach velocity profiles. When a damper is installed close to the outlet of a fan the approach velocity profile is nonuniform and much higher pressure losses through the damper can result. Figure 8.7 lists multipliers that should be applied to the damper manufacturer's catalog pressure drop when the damper is installed at the outlet of a centrifugal fan. These multipliers should be applied to all types of fan outlet dampers.

OPPOSED-BLADE DAMPER ILLUSTRATING NON-DIVERTED AIRFLOW

Figure 8.6 - Parallel Blade vs. Opposed Blade Damper

35

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AMCA 201-02 (R2011)

VOLUME CONTROL DAMPER

BLAST AREA OUTLET AREA

PRESSURE DROP MULTIPLIER

0.4

7.5

0.5

4.8

0.6

3.3

0.7

2.4

0.8

1.9

0.9

1.5

1.0

1.2

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

36

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AMCA 201-02 (R2011)

8.6 Duct branches Standard procedures for the design of duct systems are based on the assumption of uniform airflow profiles in the system.

In Figure 8.8 branch takeoffs or splits are located close to the fan outlet. Non-uniform airflow conditions will exist and pressure loss and airflow may vary widely from the design intent. Wherever possible a length of straight duct should be installed between the fan outlet and any split or branch takeoff.

Note: Avoid location of split or duct branch close to fan discharge. Provide a straight section of duct to allow for air diffusion. Figure 8.8 - Branches Located Too Close to Fan

37

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AMCA 201-02 (R2011)

loss of energy, or even a flat flange (e) on the end of the duct or fan will reduce the loss to about one half of the loss through an un-flanged entry.

9. Inlet System Effect Factors Fan performance can be greatly affected by nonuniform or swirling inlet flow. Fan rating and catalog performance is typically obtained with unobstructed inlet flow. Any disruption to the inlet airflow will reduce a fan’s performance. Restricted fan inlets located close to walls, obstructions or restrictions caused by a plenum or cabinet will also decrease the performance of a fan. The fan performance loss due to inlet airflow disruption must be considered as a System Effect.

ANSI/AMCA 210 limits an inlet duct to a crosssectional area no greater than 112.5% or less than 92.5% of the fan inlet area. The slope of transition elements is limited to 15° converging and 7° diverging.

9.2 Inlet duct elbows Non-uniform airflow into a fan inlet is a common cause of deficient fan performance. An elbow located at, or in close proximity to the fan inlet will not allow the air to enter the impeller uniformly. The result is less than cataloged air performance.

9.1 Inlet ducts Fans intended primarily for use as "exhausters" may be tested with an inlet duct in place, or with a special bell-mouthed inlet to simulate the effect of a duct. Figure 9.1 illustrates variations in inlet airflow that will occur. The ducted inlet condition is shown as (a), and the effect of the bell-mouth inlet as (b).

A word of caution is required with the use of inlet elbows in close proximity to fan inlets. Other than the incurred System Effect Factor, instability in fan operation may occur as evidenced by an increase in pressure fluctuations and sound power level. Fan instability, for any reason, may result in serious structural damage to the fan. Axial fan instabilities were experienced in some configurations tested with inlet elbows in close proximity to the fan inlet. Pressure fluctuations approached ten (10) times the magnitude of fluctuations of the same fan with good inlet and outlet conditions. It is strongly advised that inlet elbows be installed a minimum of three (3) diameters away from any axial or centrifugal fan inlet.

Fans that do not have smooth entries (c), and are installed without ducts, exhibit airflow characteristics similar to a sharp edged orifice that develops a vena contracta. A reduction in airflow area is caused by the vena contracta and the following rapid expansion causes a loss that should be considered as a System Effect. If it is not practical to include such a smooth entry, a converging taper (d) will substantially diminish the

a.

c.

b.

IDEAL SMOOTH ENTRY TO DUCT ON A DUCT SYSTEM

BELL MOUTH INLET PRODUCES FULL FLOW INTO FAN

d. CONVERGING TAPERED ENTRY INTO FAN OR DUCT SYSTEM

VENA CONTRACTA AT INLET REDUCES EFFECTIVE FAN INLET AREA

e. FLANGED ENTRY INTO FAN OR DUCT SYTEM

Figure 9.1 Typical Inlet Connections for Centrifugal and Axial Fans 38

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AMCA 201-02 (R2011)

9.2.1 Axial fans - inlet duct elbows. The System Effect Curves shown in Figure 9.2 for tubeaxial and vaneaxial fans are the result of tests run with two and four piece mitered inlet elbows at or in close proximity to the fan inlets. Other variables tested included hubto-tip (H/T) ratio and blade solidity. The number of blades did not have a significant affect on the inlet elbow SEF.

are listed on Figure 9.4, and the System Effect Curves for various square duct elbows of given radius/diameter ratios are listed on Figure 9.5. The SEF for a particular elbow is found in Figure 7.1 at the intersection of the average fan inlet velocity and the tabulated System Effect Curve. This pressure loss should be added to the friction and dynamic losses already determined for the particular elbow. Note that when duct turning vanes and/or a suitable length of duct is used (three to eight diameters long, depending on velocities) between the fan inlet and the elbow, the SEF is not as great. These improvements help maintain uniform airflow

9.2.2 Centrifugal fans - inlet duct elbows. Nonuniform airflow into a fan inlet, Figure 9.3A, is a common cause of deficient fan performance. The System Effect Curves for mitered 90° round section elbows of given radius/diameter (R/D) ratios

TUBEAXIAL FAN SHOWN

DUCT LENGTH

DUCT LENGTH

VANEAXIAL FAN SHOWN H/T

90° Elbow

Tubeaxial Fan

.25

2 piece

U

Tubeaxial Fan

.25

4 piece

Tubeaxial Fan

.35

Vaneaxial Fan Vaneaxial Fan

[1][2]

[1][2]

1.0D [1][2]

3.0D

V

W

---

X

---

---

---

2 piece

V

W

X

.61

2 piece

Q-R

Q-R

S-T

T-U

.61

4 piece

W

W-X

---

---

No Duct

0.5D

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 39

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AMCA 201-02 (R2011)

into the fan inlet and thereby approach the airflow conditions of the laboratory test setup. Occasionally, where space is limited, the inlet duct will be mounted directly to the fan inlet as shown on Figure 9.3B. The many possible variations in the width and depth of a duct influence the reduction in performance to varying degrees and makes it impossible to establish reliable SEF. Note: Capacity losses as high as 45% have been observed in poorly designed inlets such as in Figure 9.3B. This inlet condition should be AVOIDED. Existing installations can be improved with guide vanes or the conversion to square or mitered elbows with guide vanes, but a better alternative would be a specially designed inlet box similar to that shown in Figure 9.6. 9.2.3 Inlet boxes. Inlet boxes are added to centrifugal and axial fans instead of elbows in order to provide more predictable inlet conditions and to maintain stable fan performance. They may also be used to protect fan bearings from high temperature, or corrosive / erosive gases. The fan manufacturer should include the effect of any inlet box on the fan performance, and when evaluating a proposal it should be established that an appropriate loss has been incorporated in the fan rating. Should this information not be available from the manufacturer, refer to Section 10.4 for an approximate System Effect.

A counter-rotating vortex at the inlet may result in a slight increase in the pressure-volume curve but the power will increase substantially. There are occasions, with counter-rotating swirl, when the loss of performance is accompanied by a surging airflow. In these cases, the surging may be more objectionable than the performance change. Inlet spin may arise from a great variety of approach conditions and sometimes the cause is not obvious.

D

LENGTH OF DUCT

R

Figure 9.3A - Non-Uniform Airflow Into a Fan Inlet Induced by a 90°, 3-Piece Section Elbow-No Turning Vanes

9.3 Inlet vortex (spin or swirl) Another major cause of reduced performance is an inlet duct design or fan installation that produces a vortex or spin in the airstream entering a fan inlet. An example of this condition is illustrated in Figure 9.7. An ideal inlet condition allows the air to enter uniformly without spin in either direction. A spin in the same direction as the impeller rotation (pre-rotation) reduces the pressure- volume curve by an amount dependent upon the intensity of the vortex. The effect is similar to the change in the pressure-volume curve achieved by variable inlet vanes installed in a fan inlet; the vanes induce a controlled spin in the direction of impeller rotation, reducing the airflow, pressure and power (see Section 10.6).

40

Figure 9.3B - Non-Uniform Airflow Induced Into Fan Inlet by a Rectangular Inlet Duct

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AMCA 201-02 (R2011)

SYSTEM EFFECT CURVES

LENGTH OF DUCT

D

R/D

NO DUCT

—

N

+ R

2D 5D DUCT DUCT P

R-S

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

SYSTEM EFFECT CURVES

LENGTH OF DUCT

D

R/D

NO DUCT

2D 5D DUCT DUCT

0.5

O

Q

S

0.75

Q

R-S

T-U

1.0

R

S-T

U-V

2.0

R-S

T

U-V

3.0

S

T-U

V

+ R

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

LENGTH OF DUCT

D

R/D

NO DUCT

2D 5D DUCT DUCT

0.5

P-Q

R-S

T

0.75

Q-R

S

U

1.0

R

S-T

U-V

2.0

R-S

T

U-V

3.0

S-T

U

V-W

+ R

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

DETERMINE SEF BY USING FIGURE 7.1 Figure 9.4 - System Effect Curves for Various Mitered Elbows without Turing Vanes 41

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AMCA 201-02 (R2011)

SYSTEM EFFECT CURVES H R/D

NO DUCT

2D 5D DUCT DUCT

0.5

O

Q

S

0.75

P

R

S-T

1.0

R

S-T

U-V

1.0

S

T-U

V

H

LENGTH OF DUCT + R

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

H

SYSTEM EFFECT CURVES R/D

NO DUCT

0.5

S

T-U

V

1.0

T

U-V

W

2.0

V

V-W

W-X

H

LENGTH OF DUCT +

R

2D 5D DUCT DUCT

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

SYSTEM EFFECT CURVES

H

LENGTH OF DUCT

R/D

NO DUCT

0.5

S

T-U

V

1.0

T

U-V

W

2.0

V

V-W

W-X

H R

+

2D 5D DUCT DUCT

Figure 9.5C - Square Elbow with Inlet Transition - Short Turning Vanes D = Diameter of the inlet collar The inside area of the square duct (H x H) should be equal to the inside area of the fan inlet collar. * The maximum permissible angle of any converging element of the transition is 15°, and for a diverging element, 7°. DETERMINE SEF BY USING FIGURE 7.1 Figure 9.5 - System Effect Curves for Various Square Duct Elbows 42

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AMCA 201-02 (R2011)

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

IMPELLER ROTATION

COUNTER-ROTATING SWIRL Figure 9.7 - Example of a Forced Inlet Vortex

IMPELLER ROTATION

PRE-ROTATING SWIRL

IMPELLER ROTATION

COUNTER-ROTATING SWIRL

Figure 9.8 - Inlet Duct Connections Causing Inlet Spin 43

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airflow entering a duct elbow with turning vanes will leave the duct elbow with non-uniform airflow.

9.4 Inlet turning vanes Where space limitations prevent the use of optimum fan inlet conditions, more uniform airflow can be achieved by the use of turning vanes in the inlet elbow (see Figure 9.9). Numerous variations of turning vanes are available, from a single curved sheet metal vane to multi-bladed "airfoil" vanes. The pressure drop (loss) through these devices must be added to the system pressure losses. The amount of loss for each device is published by the manufacturer, but it should be realized that the cataloged pressure loss will be based upon uniform airflow at the entry to the elbow. If the airflow approaching the elbow is significantly non-uniform because of a disturbance farther upstream in the system, the pressure loss through the elbow will be higher than the published figure. A non-uniform

9.5 Airflow straighteners Figure 9.10 shows two airflow straighteners used in testing setups to reduce fan swirl before measuring stations. Figure 9.10A is the egg-crate straightener used in ANSI/AMCA 210; larger cell sizes made proportionately longer could be used. Figure 9.10B shows the star straightener used in the ISO standard. A single splitter sheet may be used to eliminate swirl in some cases. Straighteners are intended to reduce swirl before or after a fan or a process station. Do not install straighteners where the air profile is known to be non-uniform, the device will carry the non-uniformity further downstream.

TURNING VANES

TURNING VANES IMPELLER ROTATION CORRECTED PREROTATING SWIRL

TURNING VANES

CORRECTED COUNTERROTATING SWIRL

Figure 9.9 - Corrections for Inlet Spin

44

IMPELLER ROTATION

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0.45D

D

0.075D DUCT 0.075D

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

DUCT

DUCT

D

2D Figure 9.10B - ISO 5801 Star Straightener

Figure 9.10 - Test Standard Airflow Straighteners

45

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half impeller diameter between an enclosure wall and the fan inlet. Adjacent inlets of multiple double width centrifugal fans located in a common enclosure should be at least one impeller diameter apart if optimum performance is to be expected. Figure 9.11 illustrates fans with restricted inlets and their applicable System Effect Curves.

9.6 Enclosures (plenum and cabinet effects) Fans within plenums and cabinets or next to walls should be located so that air may flow unobstructed into the inlets. Fan performance is reduced if the space between the fan inlet and the enclosure is too restrictive. It is common practice to allow at least one-

2L

L EQUAL

INLET DIA.

L

EQUAL

DIAMETER OF INLET

Figure 9.11A - Fans and Plenum

L

L

Figure 9.11B - Axial Fan Near Wall

L

DWDI

L

SWSI

Figure 9.11C - Centrifugal Fan Near Wall(s)

Figure 9.11D - DWDI Fan Near Wall on One Side

L - DISTANCE INLET TO WALL

For Figures 9.11A, B & C SYSTEM EFFECT CURVES

0.75 x DIA. OF INLET 0.50 x DIA. OF INLET 0.40 x DIA. OF INLET 0.30 x DIA. OF INLET

V-W U T S

For Figures 9.11D SYSTEM EFFECT CURVES

X V-W V-W U

Determine SEF by calculating inlet velocity and using Figure 7.1

Figure 9.11 - System Effect Curves for Fans Located in Plenums and Cabinet Enclosures and for Various Wall-to-inlet Dimensions 46

L

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AMCA 201-02 (R2011)

The manner in which the air stream enters an enclosure in relation to the fan inlets also affects fan performance. Plenum or enclosure inlets or walls that are not symmetrical with the fan inlets will cause uneven airflow and/or inlet spin. Figure 9.12A illustrates this condition that must be avoided to achieve maximum performance from a fan. If this is not possible, inlet conditions can usually be improved with a splitter sheet to break up the inlet vortex as illustrated in Figure 9.12B.

common inlet obstructions. Some accessories such as fan bearings, bearing pedestals, inlet vanes, inlet dampers, drive guards and motors may also cause inlet obstruction and are discussed in more detail in Section 10. Obstruction at the fan inlet may be defined in terms of the unobstructed percentage of the inlet area. Because of the shape of the inlet cones of many fans it is sometimes difficult to establish the area of the fan inlet. Figure 9.14 illustrates the convention adopted for this purpose. Where an inlet collar is provided, the inlet area is calculated from the inside diameter of this collar. Where no collar is provided, the inlet plane is defined by the points of tangency of the fan housing side with the inlet cone radius.

For proper performance of axial fans in parallel installations minimum space of one impeller diameter should be allowed between fans, as shown in Figure 9.13. Placing fans closer together can result in erratic or uneven airflow into the fans.

The unobstructed percentage of the inlet area is calculated by projecting the profile of the obstruction on the profile of the inlet. The adjusted inlet velocity obtained is then used to enter the System Effect Curve chart and the SEF determined from the curve listed for that unobstructed percentage of the fan inlet area.

9.7 Obstructed inlets A reduction in fan performance can be expected when an obstruction to airflow is located in the fan inlet. Building structural members, columns, butterfly valves, blast gates and pipes are examples of more

SPLITTER SHEET

Figure 9.12A - Enclosure Inlet Not Symmetrical with Fan Inlet. PreRotational Vortex Induced

Figure 9.12B - Flow Condition of Figure 9.12A Improved with a Splitter Sheet. Substantial Improvement Would Be To Relocate Enclosure Inlet as Shown in Figure 9.11A

Figure 9.12 - Fan in Plenum with Non-Symmetrical Inlet

1 DIA. MIN

Figure 9.13 - Parallel Installation of Axial Flow Fans 47

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AMCA 201-02 (R2011)

ER

ET

AM

DE

SI

IN

DI

T

AR

LL

CO

E NL

I

INLET PLANE

FREE INLET AREA PLANE - FAN WITH INLET COLLAR POINT OF TANGENT WITH FAN HOUSING SIDE AND INLET CONE RADIUS

R TE E NT AM GE N DI TA OF

INLET PLANE

FREE INLET AREA PLANE - FAN WITHOUT INLET COLLAR

Table for Figure 9.14 System Effect Curve (Figure 7.1) Distance from obstruction to inlet plane Percentage of unobstructed inlet area

0.75 Inlet diameter

0.5 Inlet diameter

0.33 Inlet diameter

0.25 Inlet diameter

At Inlet plane

100

-

-

-

-

-

95

-

-

X

W

V

90

-

X

V-W

U-V

T-U

85

X

W-X

V-W

U-V

S-T

75

W-X

V

U

S-T

R-S

50

V-W

U

S-T

R-S

Q

25

U-V

T

S-T

Q-R

P

Figure 9.14 - System Effect Curves for Inlet Obstructions (Table based on Fans and Fan Systems, Thompson & Trickler, Chem Eng MAR83, p. 60)

48

10. Effects of Factory Supplied Accessories Unless the manufacturer's catalog clearly states to the contrary, it should be assumed that published fan performance data does not include the effects of any accessories supplied with the fan.

If possible, the necessary information should be obtained directly from the manufacturer. The data presented in this section are offered only as a guide in the absence of specific data from the fan manufacturer. See Figure 10.1 for terminology.

Cone Type Variable Inlet Vanes

Figure 10.1 - Common Terminology for Centrifugal Fan Appurtenances 49

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10.1 Bearing and supports in fan inlet

10.3 Belt tube in axial fan inlet or outlet

Arrangement 3 and 7 fans (see Figure 3.5) require that the fan shaft be supported by a bearing and bearing support in the fan inlet or just adjacent to it.

With a belt driven axial flow fan it is usually necessary that the fan motor be mounted outside the fan housing (see Figure 3.7 Arrangement 9, and Annex B Figure B.7).

These components may have an effect on the flow of air into the fan inlet and consequently on the fan performance, depending upon the size of the bearings and supports in relation to the fan inlet opening. The location of the bearing and support, that is, whether it is located in the actual inlet or "spaced out" from the inlet, will also have an effect. In cases where manufacturer's performance ratings do not include the effect of the bearings and supports, it will be necessary to compensate for this inlet restriction. Use the fan manufacturer's allowance for bearings in the fan inlet if possible. If no better data are available, use the procedures described in Section 9.7 as an approximation.

To protect the belts from the airstream, and also to prevent any air leakage through the fan housing, manufacturers in many cases provide a belt tube. Most manufacturers include the effects of an axial fan belt tube in their rating tables. In cases where the effect is not included, the appropriate SEF is approximated by calculating the percentage of unobstructed area of air passage way and using Figure 9.14.

10.4 Inlet box When an inlet box configuration is supplied by the fan manufacturer, the fan performance should include the effect of the inlet box.

10.2 Drive guards obstructing fan inlet All fans have moving parts that require guarding for safety in the same way as other moving machinery. Fans located less than 2.1 m (7 ft) above the floor require special consideration as specified in the United States’ Occupational Safety and Health Act. National, federal, state and local rules, regulations, and codes should be carefully considered and followed. Arrangement 3 and 7 fans may require a belt drive guard in the area of the fan inlet. Depending on the design, the guard may be located in the plane of the inlet, along the casing side sheet, or it may be "spaced out" due to "spaced out" bearing pedestals. In any case, depending on the location of the guard, and on the inlet velocity, the fan performance may be significantly affected by this obstruction. It is desirable that a drive guard located in this position be furnished with as much opening as possible to allow maximum flow of air to the fan inlet. If available, use the fan manufacturer's allowance for drive guards obstructing the fan inlet. SEF for drive guard obstructions situated at the inlet of a fan may be approximated using Figure 9.14. Where possible, open construction on guards is recommended to allow free air passage to the fan inlet. Guards and sheaves should be designed to obstruct, as little of the fan inlet as possible and in no case should the obstruction be more than 1/3 of the fan inlet area. 50

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.

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When variable inlet vanes are supplied by the fan manufacturer, the performance should include the effects of the variable inlet vane unit.

10.6 Variable inlet vane (VIV) Variable inlet vanes are mounted on the fan inlet to maintain fan efficiency at reduced airflow. They are arranged to generate an inlet vortex (pre-rotation) that rotates in the same direction as the fan impeller. Variable inlet vanes may be of two different basic types: 1) cone type integral with the fan inlet, 2) cylindrical type add-on (Figures 10.1 and 10.2).

VANE TYPE

The System Effect of a wide-open VIV (see Figure 10.2) must be accounted for in the original fan selection. If data are not available from the fan manufacturer the following System Effect Curves should be applied in making the fan selection.

SYSTEM EFFECT CURVE (100% Open)

a) Cone type, integral b) Cylindrical type

“Q” or “R” “S”

Determine SEF by calculating inlet velocity and using Figure 7.1

FAN PERFORMANCE W/OUT VARIABLE INLET VANES

120

CONE TYPE VARIABLE INLET VANES

CYLINDRICAL TYPE VARIABLE INLET VANES

PERCENT OF SHUT-OFF PRESSURE

100

VARIABLE INLET VANES 100% OPEN

75% OPEN

80 75% OPEN

60

40 75% OPEN

20

0

20

40

60

80

100

120

PERCENT OF WIDE OPEN VOLUME

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

51

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AMCA 201-02 (R2011)

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

52

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AMCA 201-02 (R2011)

Annex B. Dual Fan Systems - Series and Parallel It is sometimes necessary to install two or more fans in systems that require higher pressures or airflow than would be attainable with a single fan. Two fans may offer a space, cost, or control advantage over a single larger fan, or it may be simply a field modification of an existing system to boost pressure or airflow.

B.1 Fans operating in series To obtain a system pressure boost, fans are often installed in series. The fans may be mounted as close as the outlet of one fan directly attached to the inlet of the next fan, or they may be placed in remote locations with considerable distance between fans. The fans must handle the same mass airflow, assuming no loss or gains between stages. The combined total pressure will then be the sum of each fan’s total pressure (Figure B.1). The velocity pressure corresponds to the air velocity at the outlet of the last fan stage. The static pressure for the combination is the total pressure minus the velocity pressure and is not the sum of the individual fan static pressures. In practice there is some reduction in airflow due to the increased air density in the later fan stage(s). There can also be significant loss of airflow due to non-uniform airflow into the inlet of the next fan. Sometimes multiple impellers are assembled in a single housing and this assembly is known as a “multi-stage” fan. This combination is seldom used in conventional ventilating and air conditioning systems but it is not uncommon in special industrial systems. It is advisable to request the fan manufacturer to review the proposed system design and make some estimate of its installed performance.

These types of systems normally have common inlet and outlet sections, or they may have individual ducts of equal resistance that join together at equal velocities. In either case, the characteristic curve is the sum of the separate airflows for a given static or total pressure (Figure B.2). The total performance of the multiple fans will be less than the theoretical sum if inlet conditions are restricted or the airflow into the inlets is not straight (see Section 9.6). Also, adding a parallel fan to an existing system without modifying the resistance (larger ducts, etc.) will result in lower than anticipated airflow due to increased system resistance. Fans that have a “positive” slope in the pressurevolume curve to the left of the peak pressure curve, typical of some axial and forward curved centrifugal fans (see Figure 4.2), can experience unstable operation under certain conditions. If fans are operated in parallel in the region of this “positive” slope, multiple operating conditions may occur. Figure B.2 illustrates the combined pressure-volume curve of two such fans operating in parallel. The closed loop to the left of the peak pressure point is the result of plotting all the possible combinations of volume airflow at each pressure. If the system curve intersects the combined volume-pressure curve in the area enclosed by the loop, more than one point of operation is possible. This may cause one of the fans to handle more of the air and could cause a motor overload if the fans are individually driven. This unbalanced airflow condition tends to reverse readily with the result that the fans will intermittently load and unload. This "pulsing" often generates noise and vibration and may cause damage to the fans, ductwork or driving motors. Aileron controls in forward curved fan outlets or dampers near the inlets or outlets may be used to correct unbalanced airflow or to eliminate pulsations or reversing operation (See Figure B.3).

B.2 Fans operating in parallel Suppliers of air handling equipment and designers of custom systems commonly incorporate two identical, in parallel fans to deliver large volumes of air while taking advantage of the space savings offered by using two smaller fans.

53

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54

200%

SERIES FAN COMBINED PRESSURE CURVE

100%

SINGLE FAN PRESSURE CURVE

PERCENT OF FAN STATIC PRESSURE

SYSTEM RESISTANCE

100% PERCENT OF FAN AIRFLOW

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

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AMCA 201-02 (R2011)

STA BL ES YS TE M

UN STA BL ES YS TE M

PERCENT OF FAN STATIC PRESSURE

FAN OPERATION NOT RECOMMENDED IN THIS RANGE

100 PARALLEL FANS - FAN PRESSURE AT COMBINED VOLUME

SINGLE FAN PRESSURE CURVE

200 PERCENT OF FAN AIRFLOW

Figure B.2 - Parallel Fan Operation

AILERON

Figure B.3 - Aileron Control

55

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AMCA 201-02 (R2011)

Annex C. Definitions and Terminology C.1 The air C.1.1 Air velocity. The velocity of an air stream is its rate of motion, expressed in m/s (fpm). The velocity at a plane (Vx) is the average velocity throughout the entire area of the plane. C.1.2 Airflow. The airflow at a plane (Qx) is the rate of airflow, expressed in m3/s (cfm) and is the product of the average velocity at the plane and the area of the plane. C.1.3 Barometric pressure. Barometric pressure (pb) is the absolute pressure exerted by the atmosphere at a location of measurement (per AMCA 99-0066). C.1.4 Pressure-static. Static pressure is the portion of the air pressure that exists by virtue of the degree of compression only. If expressed as gauge pressure, it may be negative or positive (per AMCA 99-0066). Static pressure at a specific plane (Psx) is the arithmetic average of the gauge static pressures as measured at specific points in the traverse of the plane. C.1.5 Pressure-velocity. Velocity pressure is that portion of the air pressure which exists by virtue of the rate of motion only. It is always positive (per AMCA 99-0066). Velocity pressure at a specific plane (Pvx) is the square of the arithmetic average of the square roots of the velocity pressures as measured at specific points in the traverse plane. C.1.6 Pressure-total. Total pressure is the air pressure that exists by virtue of the degree of compression and the rate of motion. It is the algebraic sum of the velocity pressure and the static pressure at a point. Thus if the air is at rest, the total pressure will equal the static pressure (per AMCA 990066). Total pressure at a specific plane (Ptx) is the algebraic sum of the static pressure and the velocity pressure at that plane. C.1.7 Standard air density. A density of 1.2 kg/m3 (0.075 lbm/ft3) corresponding approximately to air at 20°C (68°F), 101.325 kPa (29.92 in. Hg) and 50% relative humidity (per AMCA 99-0066).

56

C.1.8 Temperature. The dry-bulb temperature (td) is the air temperature measured by a dry temperature sensor. Temperatures relating to air density are usually referenced to the fan inlet. The wet-bulb temperature (tw) is the temperature measured by a temperature sensor covered by a water-moistened wick and exposed to air in motion. Readings shall be taken only under conditions that assure an air velocity of 3.6 to 10.2 m/s (700 to 2000 ft/min) over the wet-bulb and only after sufficient time has elapsed for evaporative equilibrium to be attained. Wet bulb depression is the difference between drybulb and wet-bulb temperatures (td - tw) at the same location.

C.2 The fan C.2.1 Blast area. The blast area of a centrifugal fan is the fan outlet area less the projected area of the cutoff; see Figure B.6 (per AMCA 99-0066). C.2.2 Inlet area. The fan inlet area (A1) is the gross inside area of the fan inlet (see Figure 9.14). C.2.3 Outlet area. The fan outlet area (A2) is the gross inside area of the fan outlet. C.2.4 Fan. (1) A device, which utilizes a power-drive rotating impeller for moving air or gases. The internal energy (enthalpy) increase imparted by a fan to a gas does not exceed 25 kJ/kg (10.75 BTU/lbm). (2) A device having a power-driven rotating impeller without a housing for circulating air in a room (per AMCA 99-0066). The volume airflow of a fan (Q) is the rate of airflow in m3/s (cfm) expressed at the fan inlet conditions. C.2.5 Fan impeller diameter. The fan impeller diameter is the maximum diameter measured over the impeller blades. C.2.6 Fan total pressure. Fan total Pressure (Pt) is the difference between the total pressure at the fan outlet and the total pressure at the fan inlet. Pt = Pt1 Pt2 (Algebraic). Ignoring the losses that exist between the planes of measurement and the fan, Figures C.1, C.2 and C.3 illustrate fan total pressures for three basic arrangements for fans connected to external systems.

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Where the fan inlet is open to atmospheric air or where an inlet bell, as shown in the Figure C.1 is used to simulate an inlet duct, the total pressure at the fan inlet (Pt1) is considered to be the same as the total pressure in the region near the inlet (Pta) where no energy has been imparted to the air. This is the location of "still air". The following equations apply:

Where the fan outlet is open to atmospheric air or where an outlet duct three diameters or less in length is used to simulate a fan with an outlet duct and the outlet duct is open to atmospheric air, the total pressure at the fan outlet is equal to the fan velocity pressure (Pv). The following equations apply: Pt = Pt2 - Pt1 Pt2 = Pv Pt = Pv - Pt1

Pta = 0 Pt = Pt2 - Pt1 Pt1 = Pta = 0 Pt = Pt2

PLANE 1

PLANE 2

Pt2

Pt = Pt2

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

57

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AMCA 201-02 (R2011)

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PLANE 1

PLANE 2

Pt1 Pt = Pv2 - Pt1

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

PLANE 1

Pt1

PLANE 2

Pt

Pt = Pt2 - Pt1

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

58

Pt2

C.2.7 Fan velocity pressure. Fan velocity pressure (Pv) is the pressure corresponding to the average air velocity at the fan outlet. Pv = Pv2 Assuming no change in air density or area between the plane of measurement and the fan outlet, Figure C.4 illustrates fan velocity pressure. C.2.8 Fan static pressure. The difference between the fan total pressure and the fan velocity pressure. Therefore, fan static pressure is the difference between the static pressure at a fan outlet and the total pressure at a fan inlet (per AMCA 99-0066). Ps = Pt - Pv Ignoring losses between the planes of measurement and the fan, Figure C.5 illustrates the fan static pressure for a fan with ducted inlet and outlet. Ps = Ps2 - Ps1 - Pv1 (Algebraic) Where the fan inlet is open to atmospheric air, (free inlet, ducted outlet), the fan static pressure (Ps) is equal to the static pressure at the fan outlet. Ps = Ps2

PLANE 1

Pv = Pv2

Where the fan outlet is open to atmospheric air (ducted inlet, free outlet), ignoring the SEF, the fan static pressure (Ps) is equal to the inlet static pressure (Ps1) less the inlet velocity pressure (Pv1). Ps = -Ps1 - Pv1 Ps = -(-Ps1) - Pv1 Ps = Ps1 - Pv1

C.3 The system C.3.1 Equivalent duct diameter. The diameter of a circle having the same area as another geometric shape. For a rectangular cross-section duct with width (a) and height (b), the equivalent diameter is: (4ab/π)0.5 (per AMCA 99-0066). C.3.2 Fan performance. Fan performance is a statement of the volume airflow, static or total pressure, speed and power input at a stated inlet density and may include total and static efficiencies. C.3.3 Fan performance curve. Of the many forms of fan performance curves, generally all convey information sufficient to determine fan performance as defined above. In this manual, ‘fan performance curve’ refers to the constant speed performance

PLANE 2

Pv2

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

59

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AMCA 201-02 (R2011)

curve. This is a graphical representation of static or total pressure and power input over a range of volume airflow at a stated inlet density and fan speed. It may include static or total efficiency curves. The range of volume airflow that is covered generally extends from shutoff (zero airflow) to free delivery (zero fan static pressure). The pressure curves that appear are generally referred to as the pressurevolume curves. C.3.4 Normalized fan curve. A normalized fan curve is a constant speed curve in which the fan performance values appear as percentages, with 100% airflow at free delivery, 100% fan static pressure at shutoff, and 100% power at the maximum power input point. C.3.5 Point of duty. Point of duty is a statement of air volume flow rate and static or total pressure at a stated density and is used to specify the point on the system curve at which a fan is to operate. C.3.6 Point of operation. The relative position on a fan or air curtain performance curve corresponding to a particular airflow, pressure, power and efficiency (per AMCA 99-0066).

PLANE 1

Ps1

C.3.7 Point of rating. The specified fan operating point on its characteristic curve (per AMCA 99-0066). C.3.8 System. A series of ducts, conduits, elbows, branch piping, etc., designed to guide the flow of air, gas or vapor to and from one or more locations. A fan provides the necessary energy to overcome the resistance to flow of the system and causes air or gas to flow through the system. Some components of a typical system are louvers, grills, diffusers, filters, heating and cooling coils, air pollution control devices, burner assemblies, sound attenuators, the ductwork and related fittings. C.3.9 System curve. A graphic representation of the pressure versus volume airflow characteristics of a particular system. C.3.10 System Effect Factor (SEF). A pressure loss, which recognizes the effect of fan inlet restrictions, fan outlet restrictions, or other conditions influencing fan performance when installed in the system (per AMCA 99-0066).

PLANE 2

Pv1

Ps2 Ps = Ps2 - Ps1 - Pv1 (algebraic)

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

60

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AMCA 201-02 (R2011)

HOUSING

DIVERTER CU

TO

FF

CENTER PLATE BLAST AREA DISCHARGE OUTLET AREA SIDE SHEET BACKPLATE

FF

BLADE

TO

CU

INLET

SCROLL IMPELLER FRAME RIM BEARING SUPPORT INLET COLLAR

Figure C.6 - Terminology for Centrifugal Fan Components

61

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AMCA 201-02 (R2011)

CASING

BACKPLATE RIM HUB

MOTOR GUIDE VANE

INLET

BLADE IMPELLER

INLET BELL

Figure C.7A - Tubular Centrifugal FanDirect Drive CASING

BLADE DIFFUSER HUB

MOTOR

IMPELLER CASING

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

BEARING CASING BELT TUBE BLADE

HUB

GUIDE VANE IMPELLER

Figure C.7C - Vaneaxial Fan-Belt Drive

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

62

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AMCA 201-02 (R2011)

The Ps required at the fan outlet (C) will be equal to the pressure drop at the desired airflow. Since there are no inlet obstructions and the duct near the fan outlet is the same as used in the test setup, the published fan performance can be used with no additional system effect factors applied.

Annex D. Examples of the Convertibility of Energy from Velocity Pressure to Static Pressure SI CONVERSION was done using 249 Pa = 1 in. wg, 1 m3/s = 2118 cfm, 1m/s = .00508 ft/min

D.1 Example of fan (tested with free inlet, ducted outlet) applied to a duct system The overall friction of the duct system results in a 747 Pa (3.0 in. wg) pressure drop at an airflow of 1.42 m3/s (3000 cfm). SI

I-P

A

Free inlet

0.00 Pa

(no SEF)

0.0 in. wg

B-C

Outlet with straight duct attached for two or more diameters.

0.00 Pa

(no SEF)

0.0 in. wg

(duct design)

3.0 in. wg

C-D

Duct friction at Q = 1.42 m3/s (3000 cfm).

REQUIRED FAN Ps

747.00 Pa 747.00 Pa

3.0 in. wg

Select a fan for Q = 1.42 m3/s (3000 cfm) and Ps = 747 Pa (3.0 in. wg). Use manufacturer's data for rpm (N) and power (H).

NO OBSTRUCTION AT FAN INLET

Pv = 124 Pa (0.5 in.wg)

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

(I-P) in.wg (SI) Pa 996

4

747

3 Pt

498

2

249

1

0

0

Pv

Ps

A

B C

ATMOSPHERIC PRESSURE

124 Pa (0.5 in.wg)

D

Figure D.1 - Pressure Gradients - Fan as Tested 63

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AMCA 201-02 (R2011)

D.2 Example of fan (tested with free inlet, ducted outlet), connected to a duct system and then a plenum This example includes the same duct system as described in Example C.1. However, there is a short outlet duct on the fan followed by a plenum chamber with cross-sectional area more than 10 times larger than the area of the duct. The velocity in the duct from E to F is 14.4 m/s (2830 fpm), equal to a velocity pressure of 124.5 Pa (0.5 in. wg). At point "F" the Pv is 124.5 Pa (0.5 in. wg), the Ps is 0.0 Pa (0.0 in. wg), and the Pt is 124.5 Pa (0.5 in. wg). The friction of duct will cause a gradual increase in Ps and Pt back to point E. If the duct has a uniform cross-sectional area the Pv will be constant through this part of the system. Since there is an energy loss of 49.8 Pa (0.2 in. wg) as a result of the abrupt contraction from the plenum

to the duct, the Pt requirement in the plenum is 871.15 Pa (3.5 in. wg), Pt at duct entrance = 49.8 Pa (0.2 in. wg) in contraction loss, or 921.3 Pa (3.7 in. wg) Pt. Air flowing across the plenum from D to E will have a relatively low velocity and the Pv in the plenum will be 0.0 Pa (0.0 in. wg) since the velocity is negligible. At point D, there is an abrupt expansion energy loss equal to the entire Pv in the duct discharging into the plenum. The outlet duct between the fan and the plenum is 2.5 equivalent diameters long. It is the same as used during the fan rating test. The Ps in the outlet duct (also the Ps in the plenum) is the same as the Ps as measured during the rating test. This example requires a fan to be selected for 921.30 Pa (3.7 in. wg) at 1.42 m3/s (3000 cfm). Compare this with the previous selection of 747 Pa (3.0 in. wg) Ps at 1.42 m3/s (3000 cfm).

SI C-D

Outlet duct on fan as tested

0.00 Pa

D

Pv loss (also Pt loss) as result of air velocity decrease. Ps does not change from duct to plenum at D.

0.00 Pa

E E

E-F

Contraction loss - plenum to duct

I-P (no SEF)

0.0 in. wg

49.80 Pa

(part of duct system)

0.2 in. wg

Ps energy required to create velocity at E

124.50 Pa

(part of duct system)

0.5 in. wg

Duct friction at Q = 1.42 m3/s (3000 cfm)

747.00 Pa

3.0 in. wg

921.30 Pa

3.7 in. wg

REQUIRED FAN Ps Solution:

Select a fan for Q = 1.42 m3/s (3000 cfm) and Ps = 921.30 Pa (3.7 in. wg) Use manufacturer's data for rpm (N) and power (H).

64

0.0 in. wg

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AMCA 201-02 (R2011)

2.5 DIA.

NEGLIGIBLE LOSS

Pv = 124 Pa (0.5in.wg)

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

(I-P) in.wg (SI) Pa

1046 Pa (3.7 in.wg)

1245

5

996

4

747

3

498

2

922 Pa (3.7 in.wg)

922 Pa (3.7 in.wg)

Pt 747 Pa (3.0 in.wg)

249

1

0

0 A

B C

D

E

Pv 124 Pa (0.5 in.wg)

Ps

ATMOSPHERIC PRESSURE

F

Figure D.2 - Pressure Gradients - Plenum Effect

65

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AMCA 201-02 (R2011)

D.3 Example of fan with free inlet, free outlet - fan discharges directly into plenum and then to duct system (abrupt expansion at fan outlet)

energy is lost. In these applications, the energy loss and the System Effect Factor may exceed the fan outlet velocity pressure as defined in terms of "fan outlet area".

This example is similar to the plenum effect example except the duct at the fan outlet has been omitted. The fan discharges directly into the plenum.

The SEF for fans without outlet duct was obtained as follows: GIVEN:

It may seem unreasonable that the System Effect loss at the fan outlet is greater than the defined fan outlet velocity. Fans with cutoffs must generate higher velocities at the cutoff plane (blast area) than in the outlet duct (outlet area). This higher velocity (at cutoff) is partially converted to Ps when outlet ducts are used as on fan tests. When fans with cutoffs are "bulk-headed" into plenums or discharge directly into the atmosphere as with exhausters, all the velocity

B-C

SEF (see above)

B-C

Pv loss (also Pt loss) as result of air velocity decrease. Ps does not change from duct to plenum at C

D D

D-E

contraction loss - plenum to duct

Fan

Blast Area = 0 .6 Outlet Area

Fan outlet velocity = 14.4 m/s (2830 fpm) No outlet duct System Effect Curve = R-S, (from Figure 8.3) SEF = 149.4 Pa (0.6 in. wg), (from Figure 7.1) at 14.4 m/s (2830 fpm) velocity and system curve R)

SI

I-P

149.40 Pa

0.6 in. wg

0.00 Pa

0.0 in. wg

49.80 Pa

(part of duct system)

0.2 in. wg

Ps energy required to create velocity at D

124.50 Pa

(part of duct system)

0.5 in. wg

duct friction at Q = 1.42 m3/s (3000 cfm)

747.00 Pa

(duct design)

3.0 in. wg

REQUIRED FAN Ps

1070.70 Pa

Solution: Select a fan for 1.42 m3/s (3000 cfm) Q and 1070.70 Pa (4.3 in. wg) Ps. Use manufacturer's data for rpm (N) and power (H).

66

4.3 in. wg

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AMCA 201-02 (R2011)

Pv = 124 Pa (0.5 in.wg)

(I-P) in.wg

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

149 Pa (0.6 in.wg) SEF

(SI) Pa

922 Pa (3.7 in.wg)

1245

5

996

4

747

3

498

2

872 Pa (3.5 in.wg)

Pt 747 Pa (3.0 in.wg)

249

1

0

0 A

B C

D

Pv

124 Pa (0.5 in.wg)

Ps

ATMOSPHERIC PRESSURE

E

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

67

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AMCA 201-02 (R2011)

Three SEFs are shown in this example:

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

1) System Effect Curve R (see Figure 9.5 for a 3 piece inlet elbow with R/D ratio of 1 and no duct between the elbow and the fan inlet).

This example is an exhaust system. Note the entry loss at point A. An inlet bell will reduce this loss.

2) System Effect Curve U (see Figure 9.14 for a bearing in the fan inlet which obstructs 10% of the inlet).

On the suction side of the fan, Ps will be negative, but Pv is always positive.

3) System Effect Curve R (from Figure 8.3 for a fan discharging to atmosphere with no outlet duct).

Fan Pv = 124.5 Pa (0.5 in. wg)

SI A

Entrance loss - sharp edge duct

I-P

99.60 Pa

(duct design)

0.4 in. wg

(duct design)

3.0 in. wg

A-B

Duct friction at 1.42 m3/s (3000 cfm)

747.00 Pa

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

1319.70 Pa

5.3 in. wg

REQUIRED FAN Pt

Fan Ps = fan Pt - fan Pv Fan Ps (SI) = 1319.70 Pa – 124.5 Pa = 1195.2 Pa Fan Ps (I-P) = 5.3 in. wg - 0.5 in. wg = 4.8 in wg Solution: Select a fan for 1.42 m3/s (3000 cfm) Q and 1195.2 Pa (4.8 in. wg) Ps Use manufacturer's data for rpm (N) and power (H).

68

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AMCA 201-02 (R2011)

ABRUPT DISCHARGE SEF 149 Pa (0.6 in.wg)

Pv = 124 Pa (0.5 in.wg)

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

(I-P) in.wg

ELBOW SEF 149 Pa (0.6 in.wg)

(SI) Pa

OBSTRUCTION SEF 50 Pa (0.2 in.wg)

+249 +1

ATMOSPHERIC PRESSURE 0

0

-249

-1

-498

-2

-747

-3

-996

-4

100 Pa (0.4 in.wg) Pt

149 Pa (0.6 in.wg) REQUIRED Pv

-847 Pa (-3.4 in.wg)

Ps

-996 Pa (4.0 in.wg)

224 Pa (0.9 in.wg)

-1245 -5 A

-971 Pa (3.9 in.wg)

-1171 Pa (4.7 in.wg)

B

C

D

E

-1121 Pa (4.5 in.wg)

FAN INLET

Figure D.4 - Pressure Gradients - Exhaust System

69

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AMCA 201-02 (R2011)

Annex E. References These references contain additional information related to the subject of this manual: 1. ANSI/AMCA 210-99, Laboratory Methods of Testing Fans for Aerodynamic Performance Rating, Air Movement and Control Association International, Inc., 30 West University Drive, Arlington Heights, IL, 60004-1893 U.S.A., 1999. 2. AMCA Publication 200-95, Air Systems, Air Movement and Control Association International, Inc., 30 West University Drive, Arlington Heights, IL, 60004-1893 U.S.A., 1995. 3. AMCA Publication 202-98, Troubleshooting, Air Movement and Control Association International, Inc., 30 West University Drive, Arlington Heights, IL, 60004-1893 U.S.A., 1997. 4. ASHRAE Handbook, HVAC Systems and Equipment, 1996, The American Society of Heating, Refrigerating and Air Conditioning Engineers, Inc., 1791 Tullie Circle N.E., Atlanta, GA, 30329 U.S.A., 1996, (Chapter 18 Fans). 5. Traver, D. G., System Effects on Centrifugal Fan Performance, ASHRAE Symposium Bulletin, Fan Application, Testing and Selection, The American Society of Heating, Refrigerating and Air Conditioning Engineers, Inc., 1791 Tullie Circle N.E., Atlanta, GA, 30329 U.S.A., 1971. 6. Christie, D. H., Fan Performance as Affected By Inlet Conditions, ASHRAE Transactions, Vol. 77, The American Society of Heating, Refrigerating and Air Conditioning Engineers, Inc., 1791 Tullie Circle N.E., Atlanta, GA, 30329 U.S.A., 1971. 7. Zaleski, R. H., System Effect Factors For Axial Flow Fans, AMCA Paper 2011-88, AMCA Engineering Conference, Air Movement and Control Association International, Inc., 30 West University Drive, Arlington Heights, IL, 60004-1893 U.S.A., 1988. 8. Roslyng, O., Installation Effect on Axial Flow Fan Caused Swirl and Non-Uniform Velocity Distribution, Institution of Mechanical Engineers (IMechE), 1 Birdcage Walk, London SW1H 9JJ, England, 1984. 9. Clarke, M. S., Barnhart, J. T., Bubsey, F. J., Neitzel, E., The Effects of System Connections on Fan Performance, ASHRAE RP-139 Report, The American Society of Heating, Refrigerating and Air Conditioning Engineers, Inc., 1791 Tullie Circle N.E., Atlanta, GA, 30329 U.S.A., 1978. 10. Madhaven, S., Wright, T., J. DiRe, Centrifugal Fan Performance With Distorted Inflows, The American Society of Mechanical Engineers, 345 East 47th Street, New, York, NY, 10017 U.S.A., 1983. 11. Cory, W. T. W., Fan System Effects Including Swirl and Yaw, AMCA Paper 1832-84-A5, AMCA Engineering Conference, Air Movement and Control Association International, Inc., 30 West University Drive, Arlington Heights, IL, 60004-1893 U.S.A., 1984. 12. Cory, W. T. W., Fan Performance Testing and Effects of the System, AMCA Paper 1228-82-A5, AMCA Engineering Conference, Air Movement and Control Association International, Inc., 30 West University Drive, Arlington Heights, IL, 60004-1893 U.S.A., 1984. 13. Galbraith, L.E., Discharge Diffuser Effect on Performance -Axial Fans, AMCA Paper 1950-86-A6, AMCA Engineering Conference, Air Movement and Control Association International, Inc., 30 West University Drive, Arlington Heights, IL, 60004-1893 U.S.A., 1986. 14. Industrial Ventilation –23rd Edition, American Conference of Governmental Industrial Hygienists, 1330 Kemper Meadow Drive, Cincinnati, OH 45240-1634 U.S.A., 1998. 15. Fans and Systems, John E. Thompson and C. Jack Trickler, The New York Blower Company, Chemical Engineering, March 21, 1983, pp. 48-63 16. AABC National Standards, Chapter 8, Volume Measurements, Associated Air Balance Council, 1518 K Street NW, Suite 503, Washington, DC 20005 U.S.A. 70

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AMCA 201-02 (R2011)

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