79037�

A GUIDE TO FAN SELECTION AND PERFORMANCE

1. NOTATION

Units

fan casing dimension, see Sections 4 and 7.2.2: also cross-sectional area of airway according to context

mm2

fan casing dimension, see Sections 4 and 7.2.2 m

fan casing dimension, see Sections 4 and 7.2.2 m

impeller or rotor diameter at tip, see Sketch 3.7 m

impeller or rotor diameter at hub, see Sketch 3.7 m

specific diameter defined Section 3.2

frequency Hz

blade passing frequency Hz

number of sound sources

pressure loss coefficient

fan specific sound power level dB

number of blades on impeller

sound pressure level dB

sound power level dB

base sound power level dB

rotational speed of impeller rad/s

specific speed defined in Section 3.2 rad

root mean square sound pressure Pa

dynamic pressure, see Section 3.1.2 for definition Pa

fan dynamic pressure defined in Section 3.1.2 Pa

static pressure associated with system Pa

total pressure associated with system Pa

root mean square reference sound pressure Pa

fan static pressure rise defined in Section 3.1.2 Pa

A

B

C

DT

DH

d

f

fB

J

K

Kw

k

Lp

Lw

Lw∗

N

n

p

pd

pf d,

ps

pt

p0

pf s,∆

Issued December 1979 - 83 pages

1

With Amendment A

79037�

som

Subscripts

* This definition of relative density should not be confused with the ratio of the density of water vapour to that of dry aire-times associated with the term "relative density”.

fan total pressure rise defined in Section 3.1.2 Pa

fluid static power produced by fan defined in Section 3.1.2 W

fluid total power produced by fan defined in Section 3.1.2 W

power input to fan impeller W

volume flow rate through fan m3/s

distance m

reverberation time s

mean flow velocity, m/s

volume of reverberant test enclosure m3

sound power W

width of impeller flow passage at hub, see Sketch 3.7 m

width of impeller flow passage at tip, see Sketch 3.7 m

reference sound power W

fan efficiency expressed as a percentage, defined in Section 3.1.2 per cent

relative density* re 1.2 kg/m3, i.e.

density of gas flowing through fan, nominally at inlet kg/m3

density of air at standard conditions, i.e. kg/m3 kg/m3

flow coefficient defined in Section 3.2

pressure coefficient defined in Section 3.2

relates to conditions external to system

relates to conditions at impeller hub

refers to fan inlet quantity

quantity associated with octave band or measuring position

refers to mean value of quantity

refers to maximum value of quantity

pf t,∆

Pf s,

Pf t,

PR

q

r

T

U U q A⁄=

V

W

WH

WT

W0

η

σ σ ρ ρ0⁄=

ρ

ρ0 ρ0 1.2=

φ rad1–

ψ rad2–

ext

H

i

j j th

m

max

2

79037�

refers to fan outlet quantity

refers to static pressure or quantity related to fan static pressure rise

refers to total pressure or quantity related to fan total pressure rise

relates to condition at impeller tip

o

s

t

T

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ers

ated and

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.erationSectiond

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Thefan inletmately this Item.

2. INTRODUCTION

This Item contains introductory material to assist the non-specialist with fan selection and tender apIt also presents characteristic properties of the major categories of fans in a non-dimensional foallows estimates of size, power requirements and noise emission to be made for given combinationsrate and pressure rise.

To a non-specialist, the particular meanings given to certain terms in the context of fans may be coIn addition, use is made in this Item of special non-dimensional groups for generalising the characof fans. The Item therefore first leads the user, in Section 3, through comprehensive definitions anexplanations of the conventions and terminology related to fans.

For certain applications, e.g. the handling of airborne fibres, certain types of fans are a prerequisite. Se4 identifies the salient features of the major fan categories and presents their typical characterisnon-dimensional form. It will be found, however, that for many applications a choice of fans is ava

Here, apart from cost, a fan may be selected (i) according to its performance, e.g. stability of operation,ease of control, power consumption, etc., (ii) according to its mechanical arrangement, e.g. convenienceof installation, self-cleaning blade properties, etc., or (iii), because of noise emission advantages. Usfamiliar with fans, their terminology and characteristics, may refer directly to Section 7 which, for severalfan types operating at a given duty, enables impeller powers, speeds and physical sizes to be estimcompared. Alternatively, where noise emission is the primary concern, Section 9 may be used to provideestimates and comparisons of sound power spectra from several fan types. This section also givconversion formulae and other information to enable the diverse presentations of noise data which ain manufacturers' literature to be compared.

For users less familiar with the characteristics of fans, Sections 5 and 6 are necessary preliminary readingBefore selecting a fan, it is essential to have considered stability of operation, flexibility of system opincluding growth potential, and special system requirements such as back-up/stand-by capability. 5 explains these system characteristics and requirements while Section 6 gives guidance on how they anthe characteristics of fans may be successfully integrated.

In the majority of applications, fans are driven by electric motors on whose properties Section 8 givesgeneral information. It covers, for example, the starting problems that can arise with certain fan andcombinations and outlines the methods of speed control that are used for fans.

2.1 Scope

The categories of fans dealt with are: axial flow, centrifugal flow, mixed flow and crossflow. information is limited to fans where the density of the gas varies less than ten per cent between the and outlet; typically, this corresponds to a maximum fan total pressure rise, , of Pa (approxi40 inches of water). Compressors and some high pressure ratio fans are thus outside the scope of

∆pf t , 104

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3. FAN TERMINOLOGY AND BASIC RELATIONSHIPS

The following paragraphs enlarge on the terms associated with fans that are not otherwise in commowithin internal flow fluid mechanics. Section 3.1 deals with the terms surrounding fan and systeperformance and Section 3.2 enlarges on the non-dimensional groups used to portray the properties oin this Item. Section 3.3 presents the scaling approximations that are known as the Fan Laws and S3.4 describes the mechanical conventions for fans including methods for depicting the discharge anpositions of centrifugal fans.

3.1 Performance Terminology

3.1.1 General terms

Characteristic

The characteristic of a device or system is the relationship that links important primary variables asswith its operation. The most commonly used fan characteristic is the relationship between pressurevolume flow rate for a given impeller speed. Similarly the relationship between pressure loss and flow rate is the most commonly used system characteristic. Fan pressure rise and flow rate characare often expressed non-dimensionally in terms of , the pressure coefficient, and , the flow coeThese coefficients, which are fully defined in Section 3.2, are independent of impeller speed and fan si

Operating Point

An operating point is defined as the fan pressure rise/volumetric flow rate condition where the fsystem are in a stable equilibrium. This corresponds to the condition at which the fan pressure rirate characteristic intersects the system pressure loss/flow rate characteristic.

Sketch 3.1 Definition of operating point

On a system or fan that has a variable characteristic, as may be obtained using either system or fadevices, the operating point will also vary to describe an "operating line" and, where both are indepevariable, an "operating region" will result, see Sketch 3.2.

ψ φ

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total Sketch

om the

oughairingsnamic

nce

draws inuliar to

ethods

icr a large

a andible to ar, when

of the

r, for

Sketch 3.2 Influence of fan and system controls

3.1.2 Terms related to fan pressure rise

Fan total pressure rise, , often loosely termed "fan total pressure", is defined as the rise in pressure produced by the fan between its inlet and outlet. It can be defined with an open inlet, see3.3a, or open outlet (Sketch 3.3b) or as an in-duct quantity (Sketch 3.3c). Thus for to be consistentfor the different fan inlet and outlet arrangements, variations of intake and outlet losses arising frdifferent geometries are by convention neglected.

Fan dynamic pressure, , is defined as the fluid dynamic pressure corresponding to the mean thrflow velocity, , at the fan outlet based on the total outlet area without any deductions for motors, for other bodies. Thus, for the low Mach number flow rates to which this Item applies, the fan dypressure* may be taken as

. (3.1)

Fan static pressure rise, , often loosely termed "fan static pressure", is defined as the differebetween the fan total pressure rise and the fan dynamic pressure, i.e.

. (3.2)

The fan static pressure rise is thus equal to the gauge static pressure at the fan outlet when the fanair from the atmosphere through a well shaped intake. Note that the definition of this term is pecfans and is not consistent with the normal meaning of static pressure rise. The term is derived from mof testing the performance of fans.

Special care must be taken with calculations involving since it will not correspond to the actual statpressure rise between the fan inlet and outlet unless the fan draws in directly from the atmosphere oplenum, see Sketch 3.3a. When a fan has an inlet and outlet plane of the same through-flow aredischarges directly to the atmosphere, is equal to the total pressure rise across the fan for negligdensity changes, see Sketch 3.3b. Note that when discharge terminals or intakes are coupled directlyfan, they are, by convention, regarded as an integral part of the fan and not of the system. Howevethey are remote from the fan such that the fan inlet and outlet stations are within the confines

* Strictly, dynamic pressure is the difference between total pressure and static pressure and is thus greater than . Howeve air atMach numbers below 0.2, i e. 68 m/s flow velocity at standard conditions, the difference is less than one per cent.

(a) Fan controlled, e.g. by avariable speed motor

(b) System controlled, e.g. by adamper

(c) Fan and system controlled

∆pf t,

∆pf t,

pf d,Uo

½ρU2

pf d, ½ρUo2

=

∆pf s,

pf s,∆ pf t, pf d,–∆=

∆pf s,

∆pf s,

6

79037�l equale Sketch

lowugh the thoseections,

system

f t

connecting duct work, the fan total pressure rise, , and not fan static pressure rise, , wilthe static pressure rise between the fan inlet and outlet, again for negligible density changes (se3.3c).

The relationships in Sketch 3.3 all neglect variations in the inlet losses and the velocity profiles of the fentering and leaving the fan. They also neglect density changes between inlet and outlet. Althovalues of pressure differences calculated using the above relationships may differ slightly fromobtained by integrating the local velocities and pressures measured within the inlet and outlet cross sfor most practical purposes they are compatible with the pressure rises required to match ductpressure losses which are also usually calculated making similar assumptions*.

* Such duct pressure loss calculations should nevertheless include an allowance for swirl which can be pronounced downstream oube-axialfans. With the exception of diffusers, the effect of swirl is usually to increase the pressure loss of components.

† A diffuser that is directly coupled to a fan may also be regarded as an integral part of the fan in some conventions. Here the fan dynamicpressure is that corresponding to the outlet through-flow area of the diffuser. This is illustrated in Sketch 3.3a.

(a) Free intake

,

,

.

Note: .

(b) Free outlet

,

,

because fan dynamic pressure is lost at the freedischarge, i.e.

.

Hence .

∆pf t, ∆pf s,

Integral diffuser†pt( )o

pt( )i

– ∆pf t,=

ps( )o

ps( )i

– ∆pf t , ½– ρUo2

=

ps( )o

ps( )i

– ∆pf s,=

ps( )i

pt( )i

=

pt( )o

pt( )i

– ∆pf t, =

pt( )o

pt( )o'

– ½ρUo2

=

ps( )o'

pt( )o'

pt( )o

½ρUo2

–= =

pt( )o'

pt( )i

– ∆pf s, =

7

79037�

pressure

ifice

as a

Sketch 3.3 Fan ducting arrangements

Air power , , is the power delivered by a fan to the gas and is given by

, (3.3)

where the fan pressure rise, , may be taken as either the fan static pressure rise or fan total rise, i.e.

(3.4)

or . (3.5)

Note that and not is the actual power gained by the fluid. Air static power, , is an artused in the definition of fan static efficiency, , described below.

Efficiency, , is the ratio of air power, , to the power input to the propeller, , expressedpercentage, i.e.

per cent. (3.6)

Air power and hence efficiency may be taken as either a static or total quantity, i.e.

per cent (3.7)

or per cent. (3.8)

The static and total efficiencies are related by

. (3.9)

(c) Ducted inlet and outlet

,

,

.

pt( )o

pt( )i

– ∆pf t,=

ps( )o

ps( )i

– ∆pf t, =

ps( )o

pt( )i

– ∆pf s,=

Po

Po q pf∆=

∆pf

Po s, q pf s,∆=

Po t, q pf t,∆=

Po t, Po s, Po s,ηs

η Po PR

ηPo

PR------- 100×=

ηs

Po s,PR

----------- 100×=

ηt

Po t,PR

---------- 100 ×=

ηs ηt pf s , pf t,∆⁄∆=

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anothere fan

ch

,s from aressure

rmally

mental

w ratend the

nma smal

Because is the actual power delivered to the fluid, the fan total efficiency and not static efficienthe quantity representative of the fan energy conversion efficiency. The fan static efficiency is alwathan the total efficiency but this, of course, does not mean that there is any difference in the impellerequired for a given duty; the impeller power, , obtained from Equation (3.10) is identical to that obtainedfrom Equation (3.11), i.e.

(3.10)

or . (3.11)

Nevertheless, for many systems, a fan offering a high value for may be more suitable than offering a higher but lower* . This is because many systems include a discharge such that thdynamic pressure is lost from the total pressure rise, see Sketch 3.3b. The fan static pressure rise is in sucases the total pressure difference remaining to overcome the system resistance.

On systems where the fan is ducted on both the inlet and outlet (see Sketch 3.3c), the fan total pressure rise, should be used to match the fan to the system losses. However, if such a system discharge

terminal of approximately the same through-flow area as that of the fan exit plane, a fan dynamic pwill be lost and an optimised fan static efficiency will result in a minimum impeller power, .

On closed-loop systems or systems incorporating efficient diffusers, the fan total efficiency should nobe optimised.

3.2 Dimensionless Groups

Dimensionless groups are used in this Item for characterising the properties of fans of similar fundadesign but of different scale and, to a limited extent, of different proportion. Four groups are used:

specific diameter, ,specific speed, ,flow coefficient, , andpressure coefficient, .

The flow coefficient, , has a unique definition, i.e.

. (3.12)

Thus, is proportional to the ratio of two volume flow rates: that on the numerator is the actual flothrough the fan while that in the denominator is equal to the product of the impeller tip speed aimpeller surface area circumscribed by the blade tips†.

* Such a situation is possible when a fan with a large through-flow area produces a certain total pressure rise. A fan of inheretly moreefficient design (higher ) but smaller through-flow area producing the same total pressure rise at the same volume flow rate y havea lower static efficiency. The static efficiency value is penalised by the higher fan dynamic pressure associated with thelerthrough-flow area.

† For axial-flow and mixed-flow fans, the impeller surface area is ascribed a nominal value proportional to the disc area such that .See comments later in this Section.

Po t,

PR

PR q pf s,

ηs

100---------⁄∆=

PR q pf t,

ηt

100---------⁄∆=

ηsηt ηs

ηt

∆pf t,

PR

dn

φψ

φ

φq

DT3 NWT DT⁄

-----------------------------------=

φ

WT DT⁄ 1=

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79037�ise. Thus

suited to

.

e while

ve theoubleto treat

The other groups may be defined either on a basis of fan total pressure rise or fan static pressure r

(3.13)

or ; (3.14)

is a measure of the compactness of the fan. Generally, large diameter wide impellers are best high flow rates but small narrow impellers are best suited to high resistance systems.

(3.15)

or ; (3.16)

is a measure of the impeller speed necessary to produce a fan pressure rise at a given flow rate

(3.17)

or ; (3.18)

is proportional to the ratio of two pressures: that on the numerator is the actual fan pressure risthat on the denominator is the dynamic pressure corresponding to the impeller tip speed.

Any of these groups may be expressed in terms of any other two, i.e.

, (3.19)

, (3.20)

, (3.21)

. (3.22)

It may be noted that, contrary to convention, the definitions of , and used in this Item involimpeller width to diameter ratio, . The inclusion of this term allows the performance of both dinlet centrifugal fans and cross-flow fans to be correlated in terms of specific speed without having

ds DT

WT DT⁄

q-------------------

½pf s,∆

σρ0-------------=

¼

dt DT

WT DT⁄

q-------------------

½pf t,∆

σρ0------------=

¼

d

ns Nq

WT DT⁄-------------------

½ σρ0

pf s,∆-------------

¾

=

nt Nq

WT DT⁄-------------------

½ σρ0

pf t,∆------------

¾

=

n

ψs

pf s,∆

DT2 N

2σρ0

---------------------------=

ψt

pf t,∆

DT2 N

2σρ0

---------------------------=

ψ

d1

nψ½-----------

1

n1

3⁄ φ 13⁄

--------------Ψ¼

φ½--------= = =

n1

d3φ

---------1

dψ½-----------

φ½

ψ¾--------= = =

φ1

nd3

---------ψ½

d2

-------- n2ψ

32⁄= = =

ψ1

n2d

2------------ d

4φ2 φ2

3⁄

n4

3⁄---------= = =

d n φWT DT⁄

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tiped for ad

iablesce,

h a fang ,

.

cteristico beoint of

, if the if both

these fans as special cases.

Sometimes, however, when it is desirable to illustrate the performance differences between scategories of fans, conventional definitions of the groups specific diameter, specific speed ancoefficient, that do not include the term have an advantage. This is because the term , whcorrelates the effect of different impeller geometries, tends to vary more between categories of fawithin them. By omitting the term, therefore, the difference between, for example, the flow coefficietwo different categories of fans each operating at peak efficiency are exaggerated. This approach isin Sketches 4.2 and 4.3 for illustrating the inherent characteristics of different fan types. The groups spediameter, specific speed and flow coefficient defined in this way are indicated by a prime ('), e.g.

.

The success of conventional definitions in correlating the performance of single-inlet centrifugal fain the range of being limited by mechanical and fluid flow constraints within each category of Therefore, in the definitions of , but especially and (where its influence is smaller), it is an acceptabapproximation to absorb the effect of into correlations of, say, versus for each category ofcentrifugal fan.

For axial-flow and mixed-flow fans, is analogous to some function of the impeller hub to diameter ratio but, because of design constraints, it is not an independent variable in a fan optimisparticular duty. Thus, in common with the definitions of , and , the ratio should be assignea value of unity for evaluating , and for axial and mixed-flow fans, i.e. , and .

3.3 Fan Laws

In Section 3.2, the non-dimensional groups and are assembled from a combination of vargoverning the performance of fans, i.e. , , and , and variables that reflect their performani.e. and (or ).

Thus, by assigning a particular value to and knowing the values of , and associated witand its operation, the fan pressure rise is defined. Similarly, by assigning a value to and knowin

and , the flow rate is defined. Thus a fan characteristic consisting of a function of versus can, for known values of , , and , be alternatively represented by a function of versus

The function, , representing the characteristic of a given fan may be used to represent the charaof any geometrically similar* fan. If the system characteristic is known, the operating point may alsdefined as a pair of co-ordinates, and . An operating point defined in this way is known as a "prating". Geometrically similar fans having the same point of rating (i.e. operating at the same values of and ) have properties that are related by expressions known as the "Fan Laws". For examplesubscript "1" is used to denote one fan and the subscript "2" another geometrically similar fan andoperate at the same point of rating, the following expressions apply:

* Two fans are geometrically similar when one fan and a scale factor completely defines the other.

WT DT⁄ WT DT⁄

φ ′ q

DT 3N

-------------=

WT DT⁄φ' d ′ n'

WT DT⁄ d ′ n'

WT DT⁄

d ′ n' φ' WT DT⁄d n φ d d′= n n'= φ φ'=

ψ φDT N WT σ

q ∆pf s, ∆pf t,

ψ DT N ρφ DT

N WT DT⁄ ∆pf qDT N WT DT⁄ σ ψ φ

ψ φ( )

ψ φψ

φ

11

79037�

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derived

a change fand limited

teristiche

ly every.

(3.23)

and . (3.24)

For geometrical similarity, and the expressions may be rearranged into the fo

(3.25)

and . (3.26)

The fan pressure coefficient, , and pressure rises, and , in Equations (3.24) and (3.25) mayrelate to either the fan static pressure rise or fan total pressure rise.

Equations (3.25) and (3.26) can be combined to give a relationship between the impeller powers, and , i.e.

. (3.27)

Using a similar argument regarding the governing variables of fan noise, a relationship can be between the sound power levels, and , i.e.

. (3.28)

Equation (3.28) is derived from a combination of other fan laws (Equations (3.25) and (3.26)) and thefunction used in the noise estimation procedure that relates a change in pressure rise or flow rate toin sound power level. Equation (3.28) is therefore subject to the same conditions of applicability as thelaws and the noise estimation method (given in Section 9.6). Furthermore, its use for predicting sounpower level changes within individual octave bands, resulting from speed or diameter changes, isto those changes that are sufficiently small to maintain the blade passing frequency (see Equation(9.13))in the same octave band.

Sketch 3.4 demonstrates one application of the fan laws. The pressure and volume flow rate characfor a particular fan is known at a rotational speed , and it is required to predict the characteristic for tfan at a new speed . Using Equations (3.25) and (3.26), point 1 of Sketch 3.4 is transformed, movingalong a parabolic path of constant and due to the increase in speed, to a new point 2. Similarother point on the original performance characteristic may be mapped onto the new characteristic

q1

DT 3( )1N1

WT

DT--------

1

------------------------------------------ φ1 φ2

q2

DT 3( )2N2

WT

DT--------

2

------------------------------------------= = =

pf∆( )1

DT 2( )1N1

2σ1ρ0

---------------------------------------- ψ1 ψ2

pf∆( )2

DT 2( )2N2

2σ2ρ0

----------------------------------------= = =

WT DT⁄( )1

WT DT⁄( )2

=

pf∆( )2

pf∆( )1

N2

N1------

2 DT( )2

DT( )1

---------------=

2σ2

σ1------

q2 q1

N2

N1------=

DT( )2

DT( )1

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

3

ψ ∆pf( )1

∆pf( )2

PR( )1

PR( )2

PR( )2

PR( )1

N2

N1------

3 DT( )2

DT( )1

---------------=

5σ2

σ1------

Lw( )1

Lw( )2

Lw( )2

Lw( )1

20σ2

σ1------ 50

N2

N1------ 70

DT( )2

DT( )1

---------------log10 +log10 +log10 +=

N1N2

ψ φ

12

79037�

eristic.cteristicics in ating as

nownr speed.

t is thusnge of

Sketch 3.4 Effect of change of speed on fan characteristic

In this particular case, the parabolic path linking point 1 to point 2 coincides with the system charactThis is because the effect of speed implies a fan pressure rise that is proportional to , a charathat follows most system characteristics. The variation of density also affects the fan characteristway that approximately balances the effect on most systems without changing the fan point of raillustrated in Sketch 3.5.

Sketch 3.5 Effect of change of density on fan characteristic

Sketch 3.6 however demonstrates another application of the fan laws for predicting, from the kcharacteristic of one fan, the characteristic of a similar but larger fan operating at the same impelle

The paths representing constant points of rating no longer coincide with the system characteristic. Inot possible to predict operating point 2 from operating point 1 alone; instead a knowledge of a rathe original fan characteristic is necessary

q2

13

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ted in

romhat theposition

theckwise.

withfor a

Sketch 3.6 Effect of change of size on fan characteristic

Note that the fan laws are strictly only applicable for changes that involve exact geometrical similarfor incompressible flows with constant Reynolds numbers. However, in the context of fans, the efReynolds number variations is usually small and it will often be found that fans of different sizes avwithin a manufacturer's line may, for practical purposes, be regarded as geometrically similar.

3.4 Mechanical Conventions

3.4.1 Major impeller dimensions

The major impeller dimensions referred to in this Item involve diameter and width. They are illustraSketch 3.7.

3.4.2 Direction of rotation

The standard25 method of specifying the direction of rotation for centrifugal fans is to view the fan fthe drive side and indicate whether the rotation required is to be clockwise or anti-clockwise. Note tdrive side of a centrifugal fan is considered to be the side opposite the inlet even when the actual is on the inlet side.

The standard25 method of specifying the direction of rotation for axial and mixed-flow fans is to viewfan from the discharge end and indicate whether the rotation required is to be clockwise or anti-clo

3.4.3 Discharge and motor position

As with the direction of rotation, the specification of discharge position is normally only necessarycentrifugal fans. Sketch 3.8 shows the usual method of indicating discharge position and rotation centrifugal fan25.

The position of the motor for centrifugal fans should also be specified, i.e. either directly coupled oraccording to Sketch 3.9.

14

79037�

Sketch 3.7 Major impeller dimensions

Sketch 3.8 Rotation and discharge designations

15

79037�

ed on

Sketch 3.9 Standard motor positions for centrifugal fans

Axial-flow fans may be driven either by a directly coupled motor or via vee belts to a motor mountthe outside of the casing.

16

79037�

hese

ty. For largeroughacity

in termscof theseerms ofor this

snotes

4. MAJOR CATEGORIES OF FANS

There are four main categories of fans: axial flow, centrifugal flow, mixed flow and cross flow. Tcategories are defined by the nature of the flow through the impeller blades, see Sketch 4.1.

Sketch 4.1 The major fan categories

Within each of the major fan categories, the following sub-categories are covered.

Each type of fan has a different characteristic that may, or may not, be suited for the required duexample, a fan intended to blow cooling air slowly over electronic components will have a relativelyvolume flow rate capacity but relatively low pressure gain. By contrast, a fan required to blow air tha filtration system offering a high flow resistance will have a relatively small volume flow rate capbut relatively high pressure rise.

The characteristics of pressure rise and volume flow rate can be represented non-dimensionally of , the fan pressure coefficient* , and , the volume flow rate coefficient* . Fans of the same geometriproportions, though of different size and running at different speeds, when represented in terms coefficients can be compared directly. Another advantage of representing fan characteristics in tpressure and flow coefficients is for the comparison of fans from different categories although, fpurpose, it is customary to use the more conventionally defined flow coefficient†, .

* Section 3.2 gives full definitions of these terms.† The definitions of the non-dimensional groups flow coefficient, , specific speed, , and specific diameter, , employed in Sketche

4.2 and 4.3 do not include the term . Elsewhere in this Item, however, the groups are defined with a term. See in Section 3.2 for a full explanation..

Axial Flow Centrifugal Flow Mixed Flow Cross Flow

Propeller Forward-curved J-casing

Tube-axial Radial-bladed Axial-casing S-casing

Guide-vane-axial Backward-bladed U-casing

ψ φ

φ'

φ' ns' ds'

WT DT⁄( ) W

T DT⁄( )

17

79037�

fficientth mostwith the

ssure

ng therifugalles, ues andts.

low rate

r each

snotes

Sketch 4.2 Optimum efficiency fan static pressure coefficients for various types of fans

Sketch 4.2 compares the Cordier line with the pressure coefficient and flow coefficient†, , correspondingto optimum efficiency for the categories of fans dealt with in this Item. The Cordier line15 is intended torepresent pressure coefficient and flow coefficient combinations that are produced by the most edesigns of turbomachines themselves operating at optimum efficiency. It agrees acceptably well wifan categories except the forward-curved and cross-flow types whose geometry is not compatible Cordier classes of turbomachines.

Sketch 4.2 also shows that axial-flow fans can only be designed to produce relatively low precoefficients. Centrifugal-flow fans, however, produce higher coefficients.

Within the envelope for axial-flow fans, the wide range of pressure coefficient is achieved by varyisolidity, e.g. a small number of blades gives a low pressure coefficient and vice-versa. Within the centfan envelope, the range of pressure coefficient is achieved by varying the impeller blade angi.e.backward-curved blades give the lowest pressure coefficients, radial blades give intermediate valforward-curved, the highest. In addition, forward-curved blades produce the highest flow coefficien

In the same way that and relationships can be used to denote the pressure rise and fcharacteristics of a type of fan, other non-dimensional groups, namely , the specific diameter*, and ,the specific speed*, may be used to typify the impeller diameter and operating speed relationship fofan category.

* The definitions of the non-dimensional groups flow coefficient, , specific speed, , and specific diameter, , employed in Sketche4.2 and 4.3 do not include the term . Elsewhere in this Item, however, the groups are defined with a term. See in Section 3.2 for a full explanation..

φ'

ψ φ'ds' ns'

φ' ns' ds'

WT DT⁄( ) W

T DT⁄( )

18

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nt fan

ws that,

glend designduties

ratione powerthe noise by its

tures

he flow

eneralections

etailed

ion of fan

Sketch 4.3 Diagram illustrating specific speed versus specific diameter foroptimum efficiency conditions of various fans

Sketch 4.3 illustrates the relationship between specific speed and specific diameter for the differecategories when each fan within each category runs at maximum efficiency. As with Sketch 4.2, it isnoticeable how different fan categories monopolise different regions on the graphs. The sketch shofor a given impeller diameter, axial-flow fans run at higher impeller speeds than centrifugal fans.

It should be noted that Sketches 4.2 and 4.3 do not merely display patches of characteristics from a sinfan in the vicinity of its maximum efficiency. Instead the curves in the sketches connect the aconditions of several fans in each category each operating at maximum efficiency. Detailed variations between fans in the same category allow individual fans to be optimised for different within the ranges indicated in Sketches 4.2 and 4.3.

Suitability of pressure rise and volume flow rate is not the only criterion for selecting a fan. Considemust be given to the shape of the pressure rise/flow rate characteristic, the efficiency, shape of thcharacteristic, speed, size, inlet and outlet configuration and also the power and spectrum shape of emitted from different fan types. In addition, the use of a certain type of fan may be determinedmechanical suitability, e.g. for propelling airborne particulate matter.

The following Sections, 4.1 to 4.4, comment in some detail on general performance and mechanical feaof different types of fans. Guidance on their noise characteristics is given in Section 9. Typical fanperformance characteristics are illustrated in terms of , the static pressure coefficient, and , tcoefficient. Since these coefficients are directly proportional to the fan static pressure rise*, , and thevolume flow rate, , the versus charts are merely convenient methods of giving a more gpresentation than their dimensional counterparts. Also included on the same charts in the following sare typical fan static efficiencies and typical impeller power characteristics.

It should be noted that although the characteristics illustrated in Sections 4.1 to 4.4 give an indication ofthe behaviour of each category of fan, they are not wholly general within their category due to ddesign variations.

* Note that is not necessarily equal to the change in static pressure between the inlet and outlet of a fan. For a full explanatstatic pressure rise, see Section 3.1.2

ψ φ'

ψs φ∆pf s,

∆ pf s,

q ψs φ

19

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l fans, angle

sm in ther blade

e highttled.

low rate

is theber ofhan low

otherets of

ciency.izes

ection

4.1 Axial-Flow Fans

Within the major category of axial-flow fans, there are four sub-categories: propeller fans, tube-axiacontrarotating fans and guide-vane-axial fans. Most axial-flow fans are available with many bladesettings that in some cases may be adjusted, when stationary, by slackening a clamping mechaniimpeller hub. Other, more sophisticated fans, have a variable pitch facility that can alter the impelleangles while the fan is in operation.

Changing the blade angles predominantly affects the flow coefficient, . Fans optimised to producflow coefficients are set with large blade angles that can give rise to stalling if the flow is over-throStalling is manifested by a rapid decrease in both the fan pressure rise and the efficiency as the fis reduced, see for example, Sketch 4.11.

Whilst the blade angle is the chief design variable governing flow coefficient, the impeller solidity main variable governing the fan pressure coefficient. Generally, high solidity impellers (large numblades and, or, high hub to tip diameter ratios) generate high pressure coefficients more efficiently tsolidity impellers which operate more efficiently at low pressure coefficients.

4.1.1 Propeller fan

Sketch 4.4 Propeller fan

Applications for this fan mostly involve moving air through a partition from one open space to ansuch as the supply or removal of air through the roofs or walls of buildings or through the cabinelectronic equipment. The shape of the lips of the mounting ring has a marked effect on the effiSmall sizes of these units, e.g. m, often have impellers moulded from plastic but the larger susually have impellers of pressed steel. Except for very large units, e.g. m, most are suppliedwith integral electric motors and are restricted to speeds near the synchronous pole speeds (see S8).

φ

For approximatedimensions A, B, andDTsee Section 7.2

DT 0.5<DT 1.5 >

20

79037�

ratetly in a affect

nd can

nge ofe casing.

Sketch 4.5 Pressure coefficient, flow coefficient, efficiency, and impeller powercharacteristics of a typical propeller fan

4.1.2 Tube-axial fan

Sketch 4.6 Tube-axial fan

Applications for this fan are found in many air-conditioning systems that require a high volume flowbut small pressure rise. In contrast to centrifugal flow fans, this type of fan can be installed direcduct. Design variations involving the blade geometry and number of impeller blades can significantlyboth the pressure rise and flow rate.

The swirling motion of the discharged flow dissipates energy that reduces the efficiency of the fan aincrease the pressure loss of at least the first component downstream.

Fans of this type are frequently coupled directly to electric motors and thus have a restricted raoperating speeds but some versions are driven through vee belts by an electric motor mounted on th

For approximatedimensions , and

, see Section 7.2A B

DT

21

79037�

xial-flowpartedination

ng inletce the

Sketch 4.7 Pressure coefficient, flow coefficient, efficiency, and impeller powercharacteristics of a typical tube-axial fan

4.1.3 Contrarotating fan

Sketch 4.8 Contrarotating fans

These fans are used for duties that require higher pressure rises than those obtainable from single afans. The units are normally constructed from a pair of tube-axial fans so that the swirling motion imto the flow by the upstream fan is countered by the second fan. Thus the overall efficiency of the combis higher than the efficiency of either component fan in isolation.

Although the second fan is often set with smaller blade angles than the first because of the swirliflow, the maximum pressure coefficient obtainable from the combination can be more than twimaximum pressure coefficient obtainable from either component fan in isolation.

For approximatedimensions , and

,see Section7.2

A BDT

22

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makesustionnt and

outletn moreffect an

ing but it

riven.

Sketch 4.9 Pressure coefficient, flow coefficient, efficiency and impeller powercharacteristics of typical contrarotating fans

4.1.4 Guide-vane axial fan

Sketch 4.10 Guide-vane axial fan

These fans are used for duties involving high volume flow rates with modest pressure rises. Thisthem suitable for use in high velocity air-conditioning systems and also in certain forced draft combapplications. Generally, these fans are well suited to applications where high efficiency is importaalso where swirl downstream of the fan must be controlled such as in wind tunnels.

The construction of guide-vane-axial fans is similar to tube-axial fans with the addition of inlet or guide vanes. The guide vanes improve the efficiency by countering the swirl produced by the rotor. Isophisticated versions, the angle of the rotor blades may be varied while the fan is operating to eefficient flow control.

Some versions of these fans are driven through vee belts by an electric motor mounted on the casis more usual for the impeller to be directly coupled as depicted in Sketch 4.10. The vee-belt drive howeveroffers a convenient means of speed changing. Large versions of these fans are sometimes shaft d

For approximatedimensions , and

, see Section 7.2A B

DT

23

79037�

shapes entersnd is

Sketch 4.11 Pressure coefficient, flow coefficient, efficiency and impeller powercharacteristics of a typical guide-vane-axial fan

4.2 Centrifugal-Flow Fans

Within the major category of centrifugal fans, there are three sub-categories that relate to the bladeof the impeller, namely: forward curved, radial bladed and backward bladed. In all cases the flowthe impeller axially through a hollow hub (termed "eye") from both or, more usually, one end adischarged with a radial component into the casing (termed "volute", "scroll" or "shroud").

Sketch 4.12 Centrifugal fan

4.2.1 Forward-curved centrifuga1 fan

Sketch 4.13 Forward-curved impeller

For approx-imate dimen-sions , , ,

and ,see Section 7.2

A B CDT WT

24

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parede scrollimum

e flowbsorbed

tend tois thusbecause

large

This type of fan, sometimes referred to as a "volume blower" or "multi-vane" fan, is best suiapplications requiring a high volume flow rate at low to medium pressure rises. It therefore compettube-axial and guide-vane-axial fans for some duties. Unlike axial fans, however, it cannot normmounted in-line within the ductwork although some manufacturers produce a unit enclosed in a rectcasing that can be installed in-line between rectangular ducts.

The impeller accelerates the flow to a high velocity while rotating at a speed that is usually low comwith other fans. The dynamic pressure is converted to static pressure by the diffusing action of thwhose design is therefore important to the efficiency of the fan. With well-designed scrolls, the maxstatic efficiencies obtainable are comparable with those achieved by tube-axial designs.

A noteworthy feature of this fan is the shape of the power curve. If, by chance, during operation thresistance unexpectedly drops, for example due to an inspection door being opened, the power aby the impeller will rise possibly to an extent that could overload the electric motor.

Sketch 4.14 Pressure coefficient, flow coefficient, efficiency and powercharacteristics of a typical forward-curved centrifugal fan

4.2.2 Radial-discharge centrifugal fan

Sketch 4.15 Radial-discharge impellers

The chief application of radial-discharge impellers is the handling of airborne particles. The blades be self cleaning in moderately dirty conditions and the more efficient unit with curved-heel blades often used for draught induction in boilers. The fans are capable of very high pressure rises and, of their tolerance to particulate matter, are suited to filtration duties.

The simple straight-bladed radial impeller can be made from thick steel to withstand impact from

a. curved heel blades b. straight blades

25

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s the drive system

ce foro specifypensatedcy also

te power

to suitns.

re so that

particles and may be constructed with the flat blades attached directly to a spider hub. Low efficiennoisiness restricts the general application of this type of fan.

The shape of the power curve, see Sketch 4.16, is similar to that of forward-curved impellers in that it riseto a maximum when the fan static pressure rise is zero. Care must therefore be taken to ensure thatmotor is adequately rated for contingencies, such as inadvertent filter removal, that might cause theflow resistance to be reduced.

For operation in dusty environments where deposits may build up on the blades, an allowanout-of-balance forces should be made by specifying special bearings. In such cases, it is also usual ta vee-belt drive because the reduction in performance caused by deposits on the blades may be comfor by changing the pulleys in order to increase the impeller speed. However, because the efficiendeteriorates under such conditions, it is necessary to specify an electric drive motor with an adequamargin.

Sketch 4.16 Pressure coefficient, flow coefficient, efficiency and impeller powercharacteristics of a typical radial-bladed centrifugal fan

4.2.3 Backward-bladed centrifugal fan

Sketch 4.17 Backward-bladed impeller

Depending primarily on the ratio , the pressure coefficients of these impellers can be variedhigh pressure rise/low flow rate requirements or medium pressure rise/medium flow rate applicatio

The passages between the blades diffuse more than half of the dynamic pressure into static pressu

DH DT⁄

26

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ariableection

r straight

ghness

nd thatrmally

energy losses are reduced when the remaining dynamic pressure is converted to static pressure in Efficiencies are generally high but the operating speeds are higher than those of other comparabcentrifugal fans. In order to exploit the high efficiency characteristics, the inlet can be equipped with vangle guide vanes to control the flow rate. Such guide vanes, by pre-swirling the inlet flow in the dirof the impeller, provide a more efficient means of control than throttling, see Section 6.3.2.

The impellers may be constructed with aerofoil-shaped blades, constant thickness curved blades oblades.

Sketch 4.18 Blade types used with backward-bladed impellers

In principle there are advantages of efficiency with aerofoil blades and advantages of cost and touwith straight blades. Curved blades offer a frequently-used compromise.

Two further categories of backward-bladed fans are often quoted, that of wide backward-bladed aof narrow backward-bladed. There is no firm definition of these categories but "wide" fans may nobe taken to mean those with and "narrow", those with .

Sketch 4.19 Pressure coefficient, flow coefficient, efficiency and impeller powercharacteristics of a typical backward-bladed centrifugal fan

WH DT⁄ 0.3≥ WH DT⁄ 0.3<

27

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l-flow duties

-curved

4.3 Mixed-Flow Fans

Sketch 4.20 Mixed-flow fan

Mixed-flow fans have characteristics that overlap those of axial-flow fans and those of centrifugafans. Outwardly, the most common type resembles an axial-flow fan. They are normally applied tosuited to characteristics between those of axial-flow and centrifugal-flow fans.

4.3.1 Axial-casing mixed-flow fan

Sketch 4.21 Axial-casing mixed-flow fan

These fans are frequently used when a fan characteristic approximately that of a backwardcentrifugal fan is required but the installation dictates an axial inlet and outlet configuration.

Sketch 4.22 Pressure coefficient, flow coefficient, efficiency and impeller powercharacteristics of a typical axial-casing mixed-flow fan

For approximatedimensions , and , see Section7.2

A BDT

28

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is moremestic

similarenterseller istween

ponents

of the than the

4.4 Cross-Flow Fans

Sketch 4.23 Cross-flow fan

These fans are used where the convenience of the slot-like aspect of the inlet and outlet planesimportant than efficiency and where low pressure rises are sufficient. Typical applications include dofan-assisted heaters, hand-held hair dryers, and air curtains.

The fans are also known as "transverse-flow fans" and "tangential blowers". All have an impeller in cross section to those of forward-curved centrifugal impellers but, unlike centrifugal fans, the flow the impeller through a peripheral segment rather than through the hub. Thus the width of the impnot restricted by flow considerations through the "eye" of the impeller. In practice, impeller widths beabout and may be specified.

The operation of these fans is very sensitive to the shape of the casing and is easily disturbed by complaced too close to the inlet or outlet.

4.4.1 Casing configurations of cross-flow fans

Sketch 4.24 J, S and U configurations of cross-flow fans

Cross-flow fans are available with three casing configurations that determine the relative positionsintake and discharge planes. The J casing is the most common and results in a higher performanceU casing.

For approx-imate dimen-sions , , ,

and ,see Section 7.2

A B CDT WT

0.7DT 10DT

29

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Sketch 4.25 Pressure coefficient, flow coefficient, efficiency and impeller powercharacteristics of a typical J-casing cross-flow fan

30

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me

5. SYSTEM CHARACTERISTICS AND REQUIREMENTS

It is seldom that a system can operate continually at a single operating point*. For example, filter blockagein a filtration system will markedly vary the system pressure drop while the opening and closing of disterminals in a ventilation system will significantly vary the system volume flow rate. Atmosphtemperature, humidity and pressure variations will also affect the air density.

The system resistance, which may be calculated as a static pressure loss, , or total pressure lis obtained by summing the pressure losses due to the individual components in the system. The loss due to a component can be expressed as or , where and are the static-pand total-pressure loss coefficients for the component respectively; hence

(5.1)

and . (5.2)

Values of and may be obtained from sources such as Reference 8. The term or represents an externally applied pressure drop which would arise, for example, in a boiler where sigbuoyancy effects would be expected. In such a case, the applied pressure drop would be negative. Sparticular wind directions may cause applied pressure drops (either positive or negative) to systems in buildings with intakes and outlets located in different positions.

For most systems, variations in ambient temperature, humidity and pressure, whilst having no effecvolume flow rate, have an effect on the pressure drop that is represented by Equation (5.3):

, (5.3)

where subscript 1 corresponds to a condition in which the relative density† at a particular point in the systeme.g. the fan inlet, is . Subscript 2 corresponds to a condition giving the same volumetric flow rawhere the relative density† at the same point in the system is .

5.1 Fixed Systems

For the high Reynolds number flow regime to which this Item is applicable, the pressure loss will, forpurposes, be approximately proportional to the square of the flow rate. This assumes that thereautomatic control devices interfering with the natural characteristics of the system, i.e. the loss coefficients

or remain constant. If this is true, then where a system pressure loss, or corresponding to a flow rate, , is known, the pressure loss corresponding to a different flow rategiven by

(5.4)

or . (5.5)

* Operating point is defined in Section 3.1.1.† Relative density, , as defined in this Item is the ratio of the actual flow density, , to a standard flow density, equal to 1.2 kg/m3.

It should not be confused with the special meaning of relative density, i.e. ratio of water vapour density to air density, adopted by soterminologies.

∆ps ∆pt

Ks½ρU2

Kt½ρU2

Ks Kt

ps∆ Σ½ρU2Ks ps∆( )

ext+=

pt∆ Σ½ρU2Kt pt∆( )

ext+=

Ks Kt ∆ps( )ext ∆pt( )ext

ps∆( )2

ps∆( )2 ext,–

ps∆( )1

ps∆( )1 ext,–

---------------------------------------------------pt∆( )

2pt∆( )

2 ext,–

pt∆( )1

pt∆( )1 ext,–

--------------------------------------------------σ2

σ1------= =

σ ρ ρ0

σ1σ2

Ks Kt ∆ps( )1

∆pt( )1

q1 q2

ps∆( )2

σ2

σ1------

q2

q1-----

2

ps∆( )1

ps∆( )ext

–[ ] ps∆( )ext

+=

pt∆( )2

σ2

σ1------

q2

q1-----

2

pt∆( )1

pt∆( )ext

–[ ] pt∆( )ext

+=

31

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ence

lerteristic

nm will

gh thethose. These

stics tolaxing

ce. Theriteria.may be

t

The relationships given in Equations (5.4) and (5.5) enable the system operating line to be drawn, Section 3.1.1, that describes the system pressure loss variation with volume flow rate. Note, howeveEquations (5.4) and (5.5) are valid only while the loss coefficients, and , remain unchanged - hthe term "fixed system".

5.2 Systems with Varying Loss Coefficients

In many systems, the loss coefficients vary either by design, e.g. through the use of dampers and controllabdischarge terminals, or as a consequence of operation or ageing, e.g. due to the blocking of filters or othein-duct devices. This means that the system operating line will depart from the fixed characrepresented by Equations (5.4) and (5.5). This is illustrated in Sketch 5.1.

Sketch 5.1 Characteristics of system with varying loss coefficients

The passive characteristics of a system such as depicted in Sketch 5.1 should not be regarded in isolatiofrom the requirements for the system. For example, although the loss coefficients in a filtration systevary, ideally the volume flow rate through the system should not be affected. Alternatively, althouvolume flow rate through a system with controllable terminals will clearly vary, the flow rate through terminals that are open should ideally be maintained constant by an unchanging pressure dropcontrasting requirements are illustrated in Sketch 5.2.

Sketch 5.2 Constant volume flow and constant pressure drop system requirements

Although it is possible to meet very closely such requirements, by matching the system charactericontrollable fans, it is often possible to find satisfactory solutions without recourse to controls by rethe precision with which the ideal requirements are met. This is discussed in Section 6.

5.3 System Requirements for Reliability (Back-up Systems)

The previous sections have dealt with systems and their requirements in the context of performanfollowing notes, however, consider systems that are configured in certain ways to meet reliability cDepending on the criteria set, which can frequently be governed by codes of practice, systems

Curve No.

(1) Corresponds to blocked filters/closeddampers/closed terminal,

(2) corresponds to conditions inter-mediate beween (1) and (3),

(3) corresponds to clear filters/opendampers/open terminal.

(a) Constant volume flow requirement forfiltration system.

(b) Constant pressure drop requiremenfor a short ventilation system.

For curves 1, 2 and 3 see Sketch 5.1.

Ks Kt

32

79037�n series.tems.

rating inectiondevicesss antarting

rement singleistance

ailure isting but

designed to be driven by more than one fan operating in parallel, or more than one fan operating iSections 5.3.1 and 5.3.2 describe some of the special considerations that must be given to these sys

5.3.1 Systems with fans in parallel

Where failure is unacceptable, a system may be designed to be powered by two or more fans opeparallel. In this configuration, certain flow stability problems can arise and the guidance given in S6.1 should be carefully noted. To cover the event of one fan failing, the fans must be equipped with for closing the ducts to prevent short circuiting of the flow through the failed fan. In addition, unleeffective back-flow damper is provided, it may be necessary to specify a rotor lock to avoid high scurrents from unstarted fans windmilling in the wrong direction.

5.3.2 Systems with fans in series

A system employing two fans in series with one fan operating and the other idling is a common requifor certain ventilation applications. This configuration known as "series standby" is less efficient thanfan configurations due to the pressure drop caused by windmilling of the stand-by fan. This extra resmust be taken into account when calculating the system pressure drop.

Series configurations may also be used in systems where total failure is unacceptable but partial facceptable. In this case the normal operating point will be achieved only when both fans are operaan acceptable reduced flow condition will result when one fan fails. Section 6.1 gives guidance on how thecharacteristic of a two-fan system can be calculated.

33

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oint aret eitherstem totale,ystemdensityion.

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parts ofmetimes

rtainties

cessaryeral

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6. FAN AND SYSTEM MATCHING

In each of the following sections, the fan and system characteristics used to find the operating pmostly portrayed in terms of . The term is left without a subscript because it may represena fan static pressure rise and a compatible system pressure drop, or a fan total pressure rise and sypressure drop. Guidance on these aspects is given in Section 3.1.2 The relative density, , appears becausin the majority of applications, for a fixed volume flow rate, both the pressure loss exhibited by the sand the pressure rise produced by the fan are directly proportional to the density of the flow. Thus changes arising from temperature, pressure and humidity variations may be taken into considerat

It should be noted, however, that the fan pressure rise being proportional to density is conditional ufan running at constant speed. With the type of electric motor commonly used for driving fans, variations in impeller power (which is also proportional to density) will normally lead to negligible svariations but it is nevertheless important to check that the motor is adequately rated for the maexpected flow density. Further guidance on electric motor specification is given in Section 8.

The operating point should not only be positioned at the required system flow rate but situated on the fan and system characteristics that provide stable operation. In order to achieve this, it may sobe desirable to employ more than one fan. Guidance on this is given in Section 6.1.

Section 6.2 explains why matching a fan to a fixed system requires a careful assessment of the uncesurrounding system resistance estimation.

For systems with varying loss coefficients, it is necessary to decide whether control devices are neto meet the system requirements. Section 6.3 gives guidance on this and also on the choice of the sevmeans of control that are available.

Section 6.4 gives guidance on system trouble shooting.

6.1 Multi-Fan Arrangements

Situations can arise where the system pressure drop for the required flow rate exceeds that availaa single fan. Similarly, because of availability or other constraints, situations can arise where the rsystem flow rate cannot be met by a single fan. In these cases, illustrated by Sketches 6.1a and 6.1brespectively, solutions may be found by employing more than one fan.

Sketch 6.1 System requirements incompatible with characteristics of single fans

There can be other reasons dictating the use of more than one fan, e.g.

(a) two fans may fit the available space more easily than a single larger fan,

∆ p σ⁄ ∆p

σ

34

79037�other

ntilation

of theralleln total

for the

lel areinations featureed fans,cities

systemthe fans,

p

(b) volume flow rate control using multiple fan combinations may be more economical than control techniques,

(c) fans positioned near the system inlet and near the system exhaust are often used in veapplications to avoid excess pressure in the space being served,

(d) where total failure is unacceptable.

6.1.1 Two fans in parallel

With two fans connected in parallel by symmetrical ductwork, equilibrium occurs when the value total pressure rise* from both fans is identical. Thus the combined fan characteristic for two fans in pamay be constructed from the individual fan characteristics by summing, for a series of values of fapressure rise*, the values of the volume flow rate for each fan at that pressure. In Sketch 6.2 the individualcharacteristic for the two fans (which happen to be identical) is curve A-A, whilst the characteristic two fans in parallel is the curve C-C.

However, there are complications to this simple procedure.

Sketch 6.2 Combined performance characteristic for two fans in series and parallel

In Sketch 6.3, the efficiency and pressure characteristics of each of two fans operating in paralillustrated. Their combined characteristic is also shown and is drawn by adding all the possible combof volume flow rate at different pressures. The maximum in the individual pressure characteristics, aof many fan types, causes a loop in the combined characteristic which, in the case of forward-curvis likely to occur near the condition of maximum efficiency. If the combination is operated at capacorresponding to the loop, each fan may switch from either of the two conditions that satisfy the characteristic. This cyclic switching, apart from causing noise annoyance, can cause damage to other items in the system and to the fan transmission.

* If both fans, which need not be of the same size or type, discharge into a large plenum, equilibrium will occur when each fan roducesthe same fan static pressure rise, .∆ pf s,

35

79037�

tly highif underquipped ackflowployed,tarted.

otectedtic that

g

Thees.

ation is

structed

e

b

Sketch 6.3 Unstable operation of fans in parallel

In order to avoid such problems, it is necessary to select fans with pressure rises that are sufficienfor the operating point to occur on the steep portion of the combined characteristic. However, even normal operation there are no problems of instability, starting may pose problems if the fans are not ewith control devices. Where the fans have electrical starting devices or speed controls (see Section8), theyshould be started in a progressive sequence. Further, if the fans are of an axial-flow type, the bthrough the unstarted fan may induce a stall from which it may not recover. When such fans are emtherefore, it is usual to equip one fan with a throttling device to prevent backflow while that fan is s

If forward-curved or radial-bladed fans are operated in parallel, the electrical system should be pr(see Section 8.3) against the failure of one fan. This is because such fans have a power characterisincreases as the fan flow rate increases.

6.1.2 Two fans in series

The information given in this section applies only to configurations where both fans are operatin* andwhere there is insignificant interaction between the fans†.

Fans in series all pass the same volume flow, , assuming no gains or losses in the duct system. combined fan total pressure rise, , will approach the sum of the individual fan total pressure risThe fan static pressure rise for the combination is defined by Equation (3.2) and is thus not equal to thesum of the individual fan static pressure rises. Normally the fan dynamic pressure for the combintaken to be that at the outlet of the most downstream fan.

The combined fan total pressure rise versus flow rate characteristic for two fans in series may be confrom the individual fan characteristics by summing, for a series of values of , the values of the totalpressure rise for each fan at that flow rate. In Sketch 6.2 the individual characteristic of the two fans is thcurve A-A whilst the characteristic for the two fans in series is the curve B-B.

* For stand-by systems, see Section 5.3.† This will usually be a reasonable assumption for centrifugal fans irrespective of how closely coupled they may be. However, tue-axial

flow fans produce swirl which, unless damped by flow straighteners, can persist for many duct diameters. See Section 4.1.3 oncontrarotating fans.

q∆pf t,

q

36

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the fan

ust beustableeristics

erating fanry close

ous and

n would

eristicsuld beleand the

ce,

6.2 Matching Fans to Fixed Systems

Because a fixed system has a single characteristic, the operating point is determined solely bycharacteristic as depicted in Sketch 3.1.

If it is necessary to vary the flow rate through a fixed system, some means of controlling the fan mprovided either through speed control, adjustable guide vanes, variable pitch rotor control or adjdampers on either the inlet or outlet of the fan. The operating line resulting from the variety of charactoffered by such a fan is shown in Sketch 3.2a.

Sketch 6.4 Choice of fans for required flow rate

It is normally possible for more than one type of fan or fan combination to produce a satisfactory oppoint. Sketch 6.4 illustrates the characteristics of a forward-curved centrifugal fan and an axial-flowintersecting the system characteristic to secure the same operating point. Provided there are vetolerances in either the system characteristic or the fan characteristic, the choice of fan is not obvimay be swayed by considerations of efficiency, cost, ease of installation, etc. However, if the systemcharacteristic had been predicted from loss coefficients that were underestimated, the choice of fahave been important, see Sketch 6.5.

Sketch 6.5 Choice of fan influenced by potential for error in system calculations

Sketch 6.5 illustrates that the flow rate is less susceptible to estimation error in the system charactif the operating point is situated on the steep portion of the fan characteristic. However, it shorecognised that fan characteristics, although based on test measurements*, can themselves be susceptibto error due to manufacturing tolerance and, more importantly, differences between the installation

* Derivation 12 gives three classes of tolerance, A, B, and C. Manufacturers of small composite motor driven fans, e.g. most propellerfans, usually quote to class C tolerance which specifies that the volume flow rate, , for a specified duty shall not be more than 7 per centbelow the quoted value and the fan pressure rise shall not be more than 14 per cent below the quoted value. Class B toleranusuallyadopted for other types of fans, specifies corresponding tolerances of 5 per cent and 9 per cent for flow rate and pressure rise respectively.

q

37

79037�

ustratedristic.

drawnd dependsed,

icted inelect ating line. fan

tch

test set up.

The fan and system tolerances together define a region of probable operating points. Again, as illin Sketch 6.6, a smaller tolerance on the flow rate, , results from the fan with the steeper characte

Sketch 6.6 Matching fan to tolerance requirements

Note that although the shallow and steep fan characteristics used in Sketches 6.4 to 6.6 have been attributedto forward-curved centrifugal fans and axial fans respectively, no general conclusions should beabout their relative shapes. The local gradients of fan characteristics are by no means constant angreatly on the flow rate. The sketches in Section 4 may be used as a guide but before any design is finalithe manufacturer's data should be used.

6.3 Matching Fans to Systems with Varying Loss Coefficients

In the systems with more than one characteristic described in Section 5.2, the operating line resulting froman uncontrolled fan may or may not meet the required operating conditions. The operating line depSketch 3.2b is one such example. Within certain tolerances, however, it is sometimes possible to sfan with a characteristic that intersects the system characteristics to produce an acceptable operaThis is discussed in Section 6.3.1. However where that is not practicable, it is necessary to resort tocontrols and they are discussed in Section 6.3.2.

6.3.1 Meeting system requirements with fixed characteristic fans

A requirement for constant pressure rise is illustrated by line B in Sketch 5.2. This can be approximatedby selecting a fan, or fan combination, with a flat characteristic in that range of volume flow, see Ske6.7.

Sketch 6.7 Matching constant pressure requirement

q

38

79037�

in that

cases it

atural fan

ssureconstant

uld be

angle

goriese onlynance.

Similarly, the requirement for constant volume flow rate illustrated by line A in Sketch 5.2 can beapproximated by selecting a fan or combination of fans with a steep characteristic of negative sloperange of volume flow, see Sketch 6.8.

Sketch 6.8 Matching constant flow rate requirement

The approximation to constant flow rate is not as good as that to a constant pressure and in somemay be necessary to resort to a controllable fan.

6.3.2 Meeting system requirements with control devices

Fan controls are necessary when the system requirements cannot be adequately matched to the ncharacteristic.

A controllable fan offers a flexibility of operation such that, over limited ranges of flow rate and pregain, most system requirements can be met. For example, a control characteristic that gives a truly volume flow rate with increased system resistance can be achieved, see Sketch 6.9.

Sketch 6.9 Use of fan controls to achieve constant volume flow system requirement

However, if such a control characteristic is to be maintained automatically, a stability analysis shoundertaken. The starting procedure should also be modelled.

Five types of control are commonly used: speed control, variable pitch rotor control, inlet guide-vanecontrol, bleed control and throttle control.

Speed control, which can be achieved either by mechanical or electrical means, is suitable for all cateof fans and is aerodynamically the most efficient means of flow control. For cross-flow fans, it is threliable means of flow control. Mechanical controls depend on friction and require regular mainteElectrical means of control are detailed in Section 8.2.

39

79037�

ceith fixedh

irling thanible to simple

ftenlly onlyd power

The consequences of speed control can be investigated using the fan laws given in Section 3.3.

Variable pitch rotor control of axial-flow fans provides a large range of flow control with little sacrifiof efficiency. Such fans, however, are costly and require more frequent maintenance than those wrotor blades. The normal operating range variations available with such fans are illustrated in Sketc6.10.

Inlet guide vane controls can be fitted to backward-bladed centrifugal fans. The vanes act by pre-swthe inlet flow in the direction of the impeller rotation to provide a more efficient means of flow controlthrottling. However, with radial-bladed and forward-curved centrifugal fans, although it is also possgovern the flow with inlet guide vanes, the power characteristics of these fans are such that morethrottling devices can be used efficiently, see Sketches 4.14 and 4.16.

The normal operating range variations available with such devices are illustrated in Sketch 6.11.

Bleed or bypass control is a stable means of controlling the flow into a system from any fan and is oused in laboratory test rigs. In commercial operations, however, a bleed or bypass control is normaapplied to axial-flow and backward-bladed fans because they alone have characteristics that avoiconsumption penalties as the flow is diverted from the system.

Sketch 6.10 Rotor pitch control of an axial-flow fan (arbitrary units)

Sketch 6.11 Inlet guide-vane control of a backward-bladed centrifugal fan (arbitrary units)

40

79037�excepttioned

cause, as thever,

costly

re loss

dnts

ance, ithan one

e flowrmalting pointed. Itr standard for volume

s are

Throttle control is an inexpensive and often-used means of control that can be applied to all fans cross-flow fans. Throttling devices, often referred to as "dampers", should normally be posidownstream of the fan so that the flow distortion that they inevitably cause does not affect the fan.

Throttling devices are inefficient when used with axial-flow and backward-bladed centrifugal fans betheir power characteristics are such that little power will be saved, in the normal operating rangeflow is throttled. With axial-flow fans, stalling problems may arise if the flow is over-restricted. Howewhen used in conjunction with forward-curved centrifugal fans, throttling devices offer a less alternative to other means of control without incurring power consumption penalties.

The range of control available with throttling devices can be investigated by adding a pressuproportional to to the system characteristic as illustrated in Sketch 6.12.

The constants of proportionality, to , depicted in Sketch 6.12 are related to the loss coefficients anthus the position of the throttle. Item No. 690228 gives approximate data on the variation of loss coefficiewith throttle position.

Sketch 6.12 Modification of system characteristic by throttling

6.4 System Trouble Shooting

Due to uncertainties in estimating system resistance and the effects of installation on fan performis inevitable that some systems do not meet their specifications but, just as there is often more tcause of a problem, there are often several solutions.

6.4.1 Causes of low volume flow rate

If the system resistance is underestimated and/or the fan performance overpredicted, a low volumrate will result. However, if following a check it is found that errors of estimation are within nouncertainties, there are two possible causes. Either the fan has a flat characteristic near the operaas illustrated in Sketch 6.5 or some components in the fan and system are not functioning as intendshould be noted that, because of differences between the performance of a fan as measured undetest conditions12 and that occurring in an installation, it is often difficult to apportion blame on the fanan inadequate performance. Difficulties of obtaining representative measurements of pressure andflow rate in an installation compound the uncertainties.

A frequent cause of trouble is an installation that distorts the flow entering the fan. Axial-flow fan

q2

k1 k3

41

79037�sessure.ance

ility of is most Sectionthe fanainst the

essaryniently directlyir pitchdditionaltion e same

especially sensitive to inlet distortion on which Derivation 24 gives some practical guidance. Another cauof trouble is stalling in diffusers which prevents the conversion of dynamic pressure into static preItem No. 760278 gives guidance on how to correct diffuser problems. Remedies that boost fan performare offered in Section 6.4.2.

6.4.2 Boosting fan performance

If the fan inlet flow distribution and diffuser performance are satisfactory, there remains the possibboosting the fan to increase the system volume flow rate. Where fans are driven by vee belts, thiseasily achieved by changing the pulley wheels to increase the impeller speed. The fan laws given in3.3 indicate the amount by which the speed should be changed to give the required boost to performance. The corresponding increase in rotor power should also be calculated and checked agrated power of the motor.

If the fan is driven directly by an electric motor, changing the speed is impracticable. Then it is necto increase the flow and pressure coefficients of the fan. For tube-axial fans, this is most conveachieved by adding guide vanes. Many manufacturers supply guide-vane units that may be attachedto the fan casing. For guide-vane-axial fans, the effect of changing the number of blades and theangles should be investigated using manufacturer's data. If these measures are insufficient, an abooster fan* can be considered as an alternative to an increase in fan size which the fan laws in Sec3.3may be used to calculate. For centrifugal fans, a variety of impellers may be available that run in thhousing but here manufacturer's data must be used to ascertain the effect of an impeller change.

* If the booster fan is installed near the original fan, the directions of rotation of the two impellers should be contrarotating.

42

79037�

teristics against

uitable. inoosingximatengainstn

te inletd rangesr should electrichouldessarily

bility ofnt.

loyed.ited tos mayr

om a should

factoryo, a fan,

nt. Inth to

peller.

ies

7. FAN SELECTION

In selecting the best fan, the mechanical suitability, physical dimensions and performance characof the available fans should be established. The mechanical suitability of the fans should be checkedtheir intended application. For example, for duties involving dust laden air, certain fan types are unsGuidance on this is given in Section 7.1. The primary physical criterion is that the fan can be installedthe available space - convenience of installation will, in many designs, be the deciding factor for chbetween axial and centrifugal fans. A procedure for obtaining approximate dimensions and approimpeller speeds based on manufacturers' data (Section 11.3) for the different fan types is given in Sectio7.2. The primary performance criterion is that the fan should supply the required volume flow rate athe system pressure drop at an acceptable efficiency. Selection on this basis is explained in Sectio7.3.

Guidance on noise, which can also influence selection, is given in Section 9.

These procedures are illustrated by the examples in Section 10.

7.1 Mechanical Suitability

For duties involving clean and dry air at temperatures typically between 4°C and 40°C, and at absolupressures of typically between 700 mbar and 1300 mbar, most fans from manufacturers' standarwill be mechanically suitable. Outside those ranges of temperature and pressure, the manufacturebe consulted to see if alternative materials or finishes are necessary for the fan, its bearings and itsmotor. Similarly, for duties involving corrosive, wet or inflammable gases, manufacturers' guidance sbe sought before any design is finalised. There are no simple rules to imply that one fan type is necmore suited than another to such arduous conditions. The choice will often depend on the availasufficiently robust fans and those of designs that allow convenient impeller removal and replaceme

When a fan is to be used for duties involving particulate matter, centrifugal fans are normally empStraight radial-bladed centrifugal fans with open impeller construction are in general better suhandling flows containing large particles or fibres, while the more efficient shrouded impeller typebe used for less arduous dust control duties, see Section 4.2.2. Backward-bladed impellers with straight ocurved blades may also be used in the less severe environments.

With inlet flow distortion, the loss in performance and increase in vibration will be less severe frcentrifugal fan than from an axial-flow fan. When such a performance handicap is expected, a fanbe selected that gives a generous estimated performance margin, see Section 6.2.

7.2 Approximate Physical Dimensions

Because it is usually possible for more than one type of fan, or combination of fans, to provide a satissystem operation, the choice of fan may be governed by physical constraints. Where this is not sor combination of fans, may be selected on the basis of maximum efficiency, see Section 7.3.

The physical proportions of each axial-flow and mixed-flow fan type are approximately constacentrifugal and cross-flow fans, the only governing proportion that varies significantly is the widdiameter ratio of the impeller, , which determines the proportions of the casing.

Fundamental therefore to the estimation of the physical dimensions of a fan is knowledge of the imdiameter, , and, in the case of a centrifugal or cross-flow fan, the width to diameter ratio,

By correlating values of specific diameter*, , against specific speed, , coinciding with high efficienc

* See Section 3.2 for definitions and explanation.

WT DT⁄

DT WT DT⁄

ds ns

43

79037�en theure rise.timated.s

ameterpeller

ore fans referredangement

ation

lue

tained

theor, as

h

d. Forict the

ng atdure.

manyection

the

for each fan type of known , it is possible to deduce an approximate relationship betweimpeller diameter and the impeller speed, , that is best suited to the required flow rate and pressUsing this relationship between the impeller diameter and impeller speed, the diameter can be esNotes detailing the procedure for estimating are given in Section 7.2.1 and the remaining fan dimensioncan be estimated using the typical proportions listed in Section 7.2.2.

7.2.1 Procedure for estimating impeller diameter,

The following notes describe the diameter estimation procedure for a given fan type. Following the diestimation, it may become apparent that the choice of fan type should be reviewed and/or imwidth/diameter ratio changed and/or impeller speed changed. It may become evident that two or mwould be better suited to the required duty; in that case, the fan pressure rises and volume flow ratesto in the notes and charts should be taken as those appropriate to each fan in the series or parallel arrchosen.

Step Procedure

1. Select a fan type. If there are no overriding mechanical constraints, the qualitative informgiven in Section 4 may be helpful.

2. Referring to Figure 1, decide on a value for using the graph on the far left of the chart. A vaof corresponding to the middle-range value of for the chosen fan type will result in a goodcompromise between impeller diameter and speed. However, a smaller diameter can be obby specifying a higher value for .

3. Again referring to Figure 1, it can be seen that the lower portion of the graph on the far right of chart requires a value for . For axial-flow fans and mixed-flow fans, take . Fsingle inlet centrifugal fans, any number for between approximately 0.12 and 0.65typified by the values quoted on the chart*, may be chosen. For cross-flow fans, impellers wit

ranging between 0.7 and 10 are commonly available.

Note that the selection of a wide impeller reduces the diameter but increases the speeapplications where a high pressure rise is required, considerations of speed may thus restrpractical width of the impeller.

4. Evaluate , the relative density of the flow entering the fan. For sea-level installations runnioutside ambient conditions, may be taken as unity for the purposes of this estimation proce

5. For the maximum flow rate required, , calculate the necessary fan pressure rise. For systems, the pressure rise will be conveniently related to the fan static pressure rise (see S3.1.2) but, where this is not so, the fan total pressure rise, , may be calculated and convertedto using the correlation presented in Figure 4.

6. Evaluate ..

7. Pinpoint the co-ordinate on the upper far right-hand graph of Figure 1 corresponding to theselected value of and and . The point marked A exemplifies the use ofgrid for a value of , m3/s and Pa.

For footnotes see end of procedure.

WT DT⁄N

DT

DT

nsns 1 ds⁄

ns

WT DT⁄ WT DT⁄ 1=WT DT⁄

WT DT⁄

σσ

qmax

∆pf t,∆pf s,

∆pf s , σ⁄

WT DT⁄ qmax ∆pf s , σ⁄WT DT⁄ 0.5= qmax 8 = ∆pf s , σ⁄ 500=

44

79037�

timatedr a

paralleluffice.

tches in

on pecific

nal

for antal

n the

linetion

theo the

ndeed ins the

ed of

If the impeller speed and diameter are satisfactory, the remaining physical dimensions can be esusing the procedure of Section 7.2.2. If the impeller speed is very high, this may indicate the need fochange in fan type or for more than one fan to be employed in series. If the impeller speed is low, operation of two smaller fans should be considered although a double inlet centrifugal fan may sGuidance on the use of fans in series and fans in parallel is given in Section 6.1.

7.2.2 Estimation of fan external dimensions

The table below gives approximate ranges for the dimensions , and as detailed on the SkeSection 4.

7.3 Approximate Power Requirements

It is possible to estimate the impeller power of a fan using a similar approach to that given in Secti7.2.For power estimation, the crucial variable is the fan efficiency, which is correlated in terms of the sspeed, .

The procedure given in Section 7.3.1 requires a value for the specific speed, , for which a provisioestimate should be available from the procedure in Section 7.2.1.

* For double inlet centrifugal fans, the limits of this range may be doubled.† See Section 8.

8. From the far left-hand graph of Figure 1, draw a horizontal line from the chosen and co-ordinate across to one of the two Reference Lines. The point marked B exemplifies this

versus combination selected for a forward-curved centrifugal fan. Where this horizoline intersects the appropriate Reference Line, draw a line parallel to the inclined lines ocentre graph.

9. Draw a horizontal line from point A ascertained in Step 7 until it intersects with the inclined drawn in Step 8. (This is the point marked C in the running example.) This point of interseclies directly above the scale giving which can now be read.

10. To find the corresponding impeller speed turn to Figure 2. Again pinpoint point A on the farright-hand graph using the same approach as in Step 7.

11. Using the same value of as before, locate point B on one of the two Reference Lines oncentre graph and, again following the same approach as in Step 8, draw a line parallel tinclined lines.

12. Draw a horizontal line from point A to intersect with the inclined line at point C. If the right-haReference Line was used, the intersection lies directly above the scale giving the impeller sprevolutions per minute. Alternatively, if the left-hand Reference Line was used, the scale giveimpeller speed in radians per second.

13. If the fan is directly coupled to the motor, the impeller speed will be governed by the pole spethe motor. Then the nearest pole speed must be selected† and the corresponding value for ascertained by working backwards through Figure 2. Using this revised value for , thecorresponding diameter can be found by repeating Steps 2 to 9.

For footnotes see end of procedure.

l ds⁄ ns

l ds⁄ ns

DT

ns

nsns

A B C

ns

ns

45

79037�

poweric speedcedure

the

7.3.1 Procedure for estimating impeller power,

The following notes describe the impeller power estimation procedure for a given fan type. From theestimate, it may become apparent that the choice of fan type should be reviewed and/or specifvaried. This will influence the impeller speed and diameter and will necessitate repeating the progiven in Section 7.2.1.

FanSketch

No.

Propeller 4.4 for m1.25 for m

for m0.5 for m

Tube axial 4.6 for m1.1 for m

1 to 0.4

Con-rotatng 4.8 as for tube axial 0.5 to 2.0

Guide-vane 4.10 for m1.1 for m

0.6 to 1.8

Fwd-curved 4.121.4 to 2.5 1.3 to 2.1 to

Rad-bladed 4.12 1.2 to 2.1 1.2 to 1.6 0.5 to 1.0

Back-bladed 4.121.6 to 2.0 1.5 to 2.0

to

Mixed-flow 4.21 1.1 to 1.3 1

Cross-flowJ, S and U

4.241.3 to 1.65 1.75

Step Procedure

1. Referring to the far left-hand graph of Figure 3, find the fan static efficiency, , corresponding tothe value of used for estimating the impeller diameter for the fan type selected.

2. Again from Figure 3, pinpoint the co-ordinate on the far right-hand graph corresponding to values of and . (The point marked A exemplifies the use of the graph for m3/s and Pa.)

A DT⁄ B DT⁄ C DT⁄

1.70 0.50DT – DT 0.9≤DT 0.9>

0.77 0.30DT – DT 0.9≤DT 0.9>

1.25 0.15DT – DT 1.0≤DT 1.0>

1.28 0.20DT – DT 0.9≤DT 0.9>

1.20WT DT⁄ 0.1+

1.20WT DT⁄ 0.9+

1.65WT DT⁄ 0.20+

3.33WT DT⁄ 0.35+

WT DT⁄ 0.25+

PR

ηsns

qmax ∆pf s, qmax 8 =∆pf s, 500=

46

79037�

crosstion,

awn

isust be

3. From the far left-hand graph, draw a horizontal line from the value of obtained in Step 1 ato intersect with the correct Reference Line for the fan being treated (point B). At this intersecdraw a line parallel to the inclined lines on the centre graph.

4. Draw a horizontal line from point A ascertained in Step 2 to intersect with the inclined line drin Step 3. (This is point C.) This point of intersection lies directly above the scale giving which can now be read.

5. If the efficiency is satisfactory, the procedure is complete. If, however, a higher efficiencyrequired, the consequences of changing , and/or the fan type, to achieve higher values mre-investigated through the procedure in Section 7.2.1.

ηs

PR

ns

47

79037�

ction,

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ed for dustyficiencyn to

varies,

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ly

8. GUIDANCE ON ELECTRIC MOTOR SPECIFICATION FOR FANS

Almost all fans are powered by alternating current induction motors of squirrel-cage rotor construmost of which are three-phase machines to which Sections 8.1 to 8.4 are directed. Small fans, e.g. thoseabsorbing less than 1 kW, however, are usually powered by single-phase machines which are discSection 8.5.

Motors should primarily be chosen with respect to their power output and operating speed, i.e. full-loadrating*. Normally a 20 per cent maximum power margin between the fan and motor should be allowan indirectly driven unit and a 10 per cent margin for a directly coupled unit. For fans operating inenvironments, however, more generous margins should be used to allow for performance and efdeterioration, see Section 4.2.2. The fan maximum power should be calculated with consideration give

(i) the shape of the impeller power versus flow rate characteristic so that if the operating point for example due to controls or modification of the system, adequate power is available,

(ii) contingencies such as the blocking of or inadvertent removal of filters, or the breakdown of afan in the system,

(iii) the maximum density of the flow arising from, for example, extremes of ambient temperapressure and humidity. This is in contrast to the environmental conditions that should be chin relation to the maximum allowable motor operating temperature but see Section 8.4.

Induction motors, although not directly synchronised to the supply frequency, are normally rated tospeeds fairly close to the synchronous speed. Three-phase induction motors at their full rated loads run at approximately 2 per cent slip†, i.e. 0.98 times synchronous speed, but this value for slip may befor large motors, i.e. kW, or more for small motors, i.e. kW. The corresponding rotor speedare given by /no. of poles where is the motor speed in rev/min and is the sfrequency in Hz. The number of poles is typically, 2, 4, 6 or 8. Two pole machines, however, aravailable for low power duties.

Because the speeds of induction motors are related to the supply frequency, efficient speed codifficult. However, controls employing frequency regulation are available and these and other mecontrol are discussed in Section 8.2.

8.1 Starting

In general, squirrel-cage induction motors are used wherever practicable due to their low cosefficiency and good reliability. However, whilst giving satisfactory running performance, such motopresent starting problems. The speed/torque characteristic of a typical squirrel-cage induction millustrated in Sketch 8.1.

In general, a high efficiency motor, i.e. one with low resistance windings, has a low starting torque becaat the stalled condition, the nearly purely inductive rotor impedance gives it a transformecharacteristic.

Two kinds of problems arise from low starting torques: prolonged electrical heat generation and slowresponse‡. In most applications, however, it is only the first of these that requires attention. In additio

* Motors intended for speed regulation should be selected with respect to the complete fan power versus speed characterconsultation with the manufacturers. This is because speeds less than the maximum speed will give rise to higher motor temures,see Section 8.2.

† Induction motors designed for voltage regulation speed control, see Section 8.2.2, have high resistance windings that run with typicalten per cent slip at their full-load rating.

40> 10<N 0.98 120× f×= N f

48

79037�special

edtartingus it isnd its

ction

picallyces are

ngs. Alable for wiring

emplo

associated high starting currents, which can amount to ten times the full-load current, require consideration.

The heat generation rate is directly related to the rotor slip, i.e. the difference between the actual rotor speand its synchronous speed. Clearly the longer the run-up time (which will be prolonged if electrical sdevices are used to limit the starting currents) the higher the temperatures will rise in the motor. Thessential to select a motor with consideration given to the polar moment of inertia of the fan atransmission, see Table below. The time intervals between starts must also be considered, see Se8.3.

Sketch 8.1 Torque versus speed characteristic of a squirrel-cage motor and its fan

To protect the electrical system from high starting currents, it is necessary with motors of powers tygreater than 5 kW to reduce the voltage on starting. The most common electrical starting devistar/delta starters and autotransformer* starters. Normally, only two settings, i.e. "start" and "run", areoffered by either device although autotransformers can be provided with several voltage tappistar/delta device is less costly but depends on both ends of each motor winding being made avaiconnection. If either starting device is feasible, the choice may be swayed by the cost of the extraassociated with star/delta starters versus the cost of autotransformer starters.

‡ Some fan bearings and particularly vee-belt transmissions may have high breakaway torques. This torque can be overcome byyinga wound-rotor motor whose effective resistance is varied externally through slip ring connections, see Section 8.2.2.

* An autotransformer has a primary winding that is common to a portion of the secondary winding. It is thus less costly than a conventionaltransformer whose windings are mutually isolated.

49

79037�es.

choice of

pitch

thatay be

nsulted.

elta

The following table may be used as a guide to the relative suitabilities of different starting techniqu

The above values should be taken only as a guide and it should be noted that in many cases, the starting arrangement is governed by electricity supply mandates.

Sketch 8.2 Fan motor starting characteristics

The problem of insufficient starting torque has several solutions:

(i) an aerodynamic solution in which the fan torque curve is flattened by the use of a variablefan, the temporary provision of a bypass, or temporary throttling,

(ii) the specification of a larger size of electric motor,

(iii) the use of an autotransformer starter with several voltage tappings,

(iv) the use of a wound-rotor induction motor, see Section 8.2.1,

(v) the use of a slipping coupling, for example, a fluid or powder coupling.

A frequently-used rule of thumb9 to check that starting will not cause overheating is to specify a motorhas a run-up time within those obtained from the following table. However, longer run-up times msatisfactory in many installations but where this is envisaged the motor manufacturer should be co

Power ratings Suitable starting arrangement

kW Single-phase, no external starting arrangement

kW to kW Three-phase, no external starting arrangement

kW to kW Three-phase, star/delta starter

above 30 kWup to 75 kW

Three-phase, wound rotor motor with external resistance. Used only if star/dand autotransformer are unsuitable for torque reasons

above 37.5 kW Three-phase, autotransformer starter

0 1 –

0.5 5.0

5.0 37.5

50

79037�

ried by enables

que.

The actual starting or run-up time may be calculated from the following expressions.

(8.1)

8.2 Speed Control

There are five means of electrical speed control that may be applied to fans. They are:

(i) control employing a wound-rotor induction motor,

(ii) control employing voltage regulation to an induction motor,

(iii) control employing frequency regulation to an induction motor,

(iv) control employing a two speed induction motor,

(v) control employing a direct current motor.

8.2.1 Wound-rotor induction motors

The resistance, and hence current and torque, of a wound-rotor induction motor can be vaconnections, made through slip rings, between the motor windings and external resistances. Thisthe speed to be governed by the torque versus speed characteristic of the fan, see Sketch 8.3.

Starting method Maximum allowable run-up time (in seconds)

Direct on-line

Star/delta or autotransformerwith approximately 60 per centfull voltage

Wound rotor motor for powers up to 37.5 kW

or for powers above 37.5 kW

where is the motor speed (rad/s) at full load,

is the polar inertia of the fan, drive and motor (referred to motor) in kg m2 and

iS the average torque available for acceleration divided by the full-load torTypically for star/delta starting and 1 for direct on-line starting.

3Rated Power kW( )

7.5------------------------------------------------+

9Rated Power kW( )

2.3------------------------------------------------+

8Rated Power kW( )

3------------------------------------------------+

18Rated Power kW( )

15------------------------------------------------+

TstartIN

210

3–×τ Rated Power (kW)×-------------------------------------------------------=

N

I

ττ 0.25=

51

79037�

uencyeristic is

nables

woundasted as heatuld be

s the fan speed high

Sketch 8.3 Torque/speed characteristics of wound-rotor motor

Due to the increase in slip inevitable with reducing the speed of any induction motor with a fixed freqsupply, the efficiency deteriorates as the speed is lowered. However, when the fan power characttaken into account, power savings nevertheless result.

8.2.2 Voltage regulation

Like the method employing wound-rotor motors, this is essentially a means of torque control that ethe speed to be governed by the torque/speed characteristic of the fan, see Sketch 8.4.

To alter the voltage, thyristors are often employed in the supply to the induction motor.

Sketch 8.4 Torque/speed characteristics with induction motor voltage regulation

The slip associated with such means of control leads to low efficiencies at low speeds but, as with therotor induction motor control, power savings result. Unlike the wound-rotor motor, however, the wpower, which reaches a peak at approximately 75 per cent of the maximum speed, is dissipatedwithin the motor rather than in external resistances. The provision for such heat dissipation shoarranged in consultation with the motor manufacturer. Generally, as motor sizes rated at three timemaximum power are usually specified, cost considerations rule out the application of this form ofcontrol for powers in excess of approximately 30 kW. The motors are normally equipped withresistance windings in order to widen the stable operating speed range.

52

79037�

d to theuency

enesshould be

-speeds, poleding issignedle PAM

windingcircuits.h their

usuallyom the

eds

8.2.3 Frequency regulation

Because the speed of an induction motor, when operating at its rated condition, is closely relatesupply frequency, it may remain largely at its rated condition yet vary in speed if the supply freqvaries. A typical torque/speed characteristic is illustrated in Sketch 8.5.

Sketch 8.5 Torque/speed characteristics with induction motor frequency regulation

This type of control is very much more efficient than one involving high slip. However, as the effectivof the motor cooling may decrease as the speed is reduced, the provision of adequate cooling schecked with the motor manufacturer.

The high costs of this type of control normally restrict the application to high power duties, e.g. aboveapproximately 40 kW.

8.2.4 Two-speed induction motors

In many fan applications, a limited range of speed control is sufficient. There are three types of twomotors currently in production that offer a stepped means of speed control: dual-winding motorchanging (or Dahlander) motors and pole amplitude modulation (PAM) motors. The Dahlander wina special case of the PAM winding that offers a speed ratio of 2:1 only. A PAM winding can be deto produce a speed ratio corresponding to any number of poles combination. For example, a 6 pomotor can be wound to be switched from a 6 pole speed to a 4 pole speed. The traditional dual-motor has the same flexibility but suffers from problems of space that can compromise the magnetic Efficiency is thus often a little lower than Dahlander or PAM motors operating at high speed for whicwindings are optimised. However, star/delta starting on pole changing motors* is not possible withoutspecial designs. Therefore for medium to high power requirements, e.g. above about 5 kW, dual-windingmotors are often specified.

Although two-speed motors are more expensive than single-speed motors, the additional cost will be significantly less than other means of speed control. In addition, the control does not suffer frefficiency and heat dissipation problems associated with slip.

* It may nevertheless be acceptable, provided high current transients are acceptable, to start such motors at their lower spebeforeswitching to their higher speeds.

53

79037�

blemsund, d.c. othern rated

ethodting andrticularl is, in

size, arech mustg.

hereforegs. Suchhich

ed by ther inertia to protect

which follows:

nde

Sketch 8.6 Torque/speed characteristics of a pole amplitude modulated motor

8.2.5 Controls employing direct current motors

The efficient control of d.c. motors is relatively simple compared with a.c. machines and starting proare obviated because the control range extends from the maximum to zero speed. When series womotors have a limited power consumption which makes them especially suitable for marine andinstallations where the power supply may have a limited capacity. However, d.c. machines, for a givepower, are more costly and require more maintenance than squirrel-cage induction motors.

Although d.c. motors are conventionally controlled by external resistors, an increasingly popular muses thyristor rectifiers. Here, the harmonics generated by the controls may induce extra motor heareduce brush life and, although most modern motors are designed appropriately, the suitability of a pamotor for this means of control should be checked with the manufacturer. This means of controgeneral* , less expensive than frequency regulation for motors rated below about 40 kW.

8.3 Overload Protection

Modern electric motors are rated to operate at high temperatures and, because of their compactmore susceptible to overheating than their earlier counterparts. Simple fuses or circuit breakers whiwithstand high starting currents will not in general provide adequate protection against overheatin

Whether or not overheating protection is an economic proposition largely depends on the size and tcost of the motor. Thus on larger motors temperature transducers are often embedded in the windinprotection is particularly desirable for forward-curved and radial-bladed centrifugal fan motors wabsorb more power as the system resistance decreases. Overloading may therefore be causaccidental opening of an inspection door or a runaway in an automatic control system. The high polaof large versions of these types of fans means that temperature transducers may also be necessarythe motor from overheating as a result of too frequent starting.

8.4 Motor Construction

Motors are available with various enclosures and classes of insulation to suit the environment withinthey will be operated. These are defined in the appropriate British Standards but are summarised as

* Due to the high costs of d.c. motors suitable for use in hazardous environments, frequency regulation may be more economic ur suchconditions for powers less than 40 kW.

54

79037�

sses oflife can

loadedowever,l, givenhose

videde system

00 watts.ypes ofin excess

isks of

Enclosure

Insulation

The standard ambient temperature for motor design is . Associated with this are seven clainsulation, summarised below, that classify the maximum temperature rise at which normal service be expected.

For motors operating at high altitude, the allowable temperature rise will be exceeded if the motor isto produce its rated power because the cooling effectiveness is reduced by the lower air density. Hif the fan maximum power as ascertained in note (iii) in Section 8 is related to the flow density at sea levethe motor will not overheat at any higher elevation because the impeller power requirement, for apoint of rating* decreases faster than the motor capability. Similarly, a motor selected for a fan wmaximum power is related to the flow density at will not overheat at lower temperatures prothe motor is also subject to the lower temperatures. This is a useful feature when a high temperaturis started cold.

8.5 Single-Phase Motors

Single-phase squirrel-cage motors are almost always used for fan applications requiring less than 5The properties of the most frequently used types are summarised in the following table. Other tsingle-phase motors include a.c. series motors (universal motors) which are used where speeds

* See Section 3.3 for definition.

Screen-protected – for use in relatively clean and dry locations,

Drip-proof – for use as above, but where liquid is likely to drip onto the motor,

Totally-enclosed – for use in areas where dirt, water or vapour makes enclosure desirable,

Flameproof – for use in a hazardous area where there are inflammable gases or rexplosion,

Weatherproof – for use without further protection from weather conditions.

Class Temperature rise above 40 °C

Y

A

E – in normal use

B – in normal use

F – in normal use

H

C

40°C

50°C

65°C

80°C

90°C

115°C

140°C

140> °C

40°C

55

79037�h have

of the maximum a.c. synchronous speeds are required and split-phase induction motors whic

characteristics somewhat similar to capacitor-start motors.

Single-Phase Motors for Fans

Motor type

DescriptionRated output(W)

Ratedspeeds50 Hz(rev/min)

Ratedspeeds60 Hz(rev/min)

Efficiency (per cent)

Startingtorque(per cent

of full load)

Applications

AdvantagesDisadvantag

es

Shaded

pole

induction

motor

Usually of open

construction,

starting torque is

provided by

permanently

short-circuited

auxiliary winding.

0.75

to

200

875

1300

2600

1050

1550

3100

up

to

40

30

to

80

Low power

fans, e.g.

propeller and

cross-flow

requiring no

maintenance.

Inexpensive,

multispeed

capability,

quite small.

Low efficiency,

low starting

torque.

Capacitor

start,

induction

run

Usually of

enclosed

construction, has

auxiliary starting

winding in series

with external

capacitor. This is

isolated by

centrifugal switch

as normal running

speed is

approached.

40

to

800

950

1425

2850

1140

1725

3450

35

to

50

165

to

240

Fans in small

commercial

equipment,

suited to belt

drives owing

to high

starting

torque.

High starting

torque.

Non-adjustable

speed.

Capacitor

start,

capacitor

run

Of enclosed

construction, has

two windings, one

that is 90°

electrically

phased from other

by capacitor.

20

to

2000

900

1350

2700

1075

1625

3250

45

to

60

30

to

250

(high value

only with

additional

starting

capacitor

with

centrifugal

switch)

Medium size

directly

driven fans.

High starting

torques are

possible. Can

be of high

efficiency and

fitted with

speed and

reversal

controls.

Speed control is

not possible

with

two-capacitor

versions.

Single capacitor

versions have

low starting

torque and load

sensitive speed.

56

79037�

epends on noise

nerationgy and general

neratedre likely

ulent airhe ductic forcesinto the ambient

gh whiche noisetsigned,irbornee use ofDirecte use of

s of afor

r somenvenientnitudesase 10 power

ely. Thes media

9. GUIDANCE ON NOISE

9.1 Introduction

There is an increasing emphasis placed on acceptable noise levels so that, because the noise dthe system configuration, the design of duct systems should not be undertaken in isolation fromconsiderations. To enable duct system designers to understand more fully the principles of noise geand propagation within duct systems this Section provides an introduction to acoustic terminolorelationships used in fan noise estimation. A means of estimating the fan sound power level, andguidance on the selection and operation of fans for minimum noise, is also provided.

The noise within a duct system can originate from several sources other than the fan. Noise will be geby the fan drive, bearings and gearing but for other than very low speed fans these noise sources ato be dominated by aerodynamic noise generated by the fan. Noise will also be generated by turbflows resulting from obstructions, such as bends and dampers within the duct, and branching of tsystem; additional noise may be generated due to structural resonance effects induced by periodassociated with the blade passing frequency or vortex shedding. Noise may also be introduced system from external sources; for example, when the duct system passes through an area with highnoise levels.

The noise emitted from the system may be transmitted to the serviced areas, or other areas throuthe ducting system passes, through the air, the system structure or directly via the duct. Airbornprincipally emanates from the plant room and can be attenuated (i.e. reduced) by silencing the ducdownstream of the fan and acoustically insulating the plant room. However, if the system is poorly dethe flow within the duct may cause the duct panels to vibrate so providing an additional source of anoise. Structure borne noise may be reduced by isolating the driver and fan from the system by thanti-vibration mounts and by using flexible couplings at suitable positions in the duct system. ductborne noise is introduced into the serviced area through louvres and may be reduced by thspecial acoustic louvres or by noise suppressors between the fan and ventilators.

Although knowledge of the fan noise alone will clearly not be sufficient to determine the noisinessystem, the fan sound power level, , is an important input to such calculations. Procedures determining the noisiness of ventilation systems are available in References 3 and 6 and a method of ratingindustrial noise and its effect on persons living in the vicinity is given in Reference 2.

9.2 Noise Units and Basic Relationships

Sound powers radiated in air have a practical range of about W to W. Sound powers fofamiliar noise sources are tabulated in this Section. Because this range of powers is so large it is coto express sound powers in decibel units (dB). This unit may be used to compare the relative magof any quantity related to power. The number of bels (1 bel = 10 decibels) is the logarithm to the bof the ratio of the powers. The two quantities used in this Item that are measured in dB are soundlevel, , and sound pressure level, . These quantities are defined as follows:

(9.1)

and , (9.2)

where and are reference sound power and root mean square sound pressure respectivpreferred* reference power, , is W and the preferred reference sound pressure in gaseou

Lw

1012–

108

Lw Lp

Lw 10 WW0-------

log10 =

Lp 20 pp0-----

log10 =

W0 p0W0 10 12–

57

79037�measure

e) levelsnds (quencye bands

is Pa*; these values are used in this Item. In air, sound pressure level is the easiest quantity to so that the noise of fan and duct systems is measured in terms of .

The quality of noise is dependent both on its volume and frequency content (i.e. its spectral composition).For a full description of a noise source, it is therefore necessary to quote sound power (or pressurwithin specific frequency bands. The frequency bands used for fan noise description are octave bai.e.bands of frequency having an upper frequency twice that of the lower frequency). The preferred frelimits for each octave band, and their nominal centre frequencies, are tabulated for the eight octavof interest.

* The preferred reference values correspond to the youthful hearing threshold and are expressed in SI units in this Item. It should be stressedthat for all quantities expressed in dB the reference value used should be stated when absolute values are given.

* 1 Pa is 1 pascal = 1 N/m2; therefore N/m2. The equivalent value of may be quoted in other units such as dyn/cm2 or lbf/ft2.

20 µ

20 µPa 25–×10= p0 2

4–×104.18

7–×10

Lp

58

79037�

, are the

uote alevel andonse, toaid to be

make aer octave octave

d bandl using

re knownressure

assessedfluencednd noise

vidualin

It should be noted that the band centre frequencies, by which the bands are usually referred tonominal geometric mean of the upper and lower band frequency limits, i.e. .

9.3 Reaction to Noise

Values of may be quoted for each octave frequency band but it is often more convenient to qsingle number. In order that the number meaningfully represents both the absolute sound pressure frequency content of the sound it is necessary to apply a correction, based on the observer respassign to individual octave bands an equivalent loudness. This corrected sound pressure level is sweighted.

Since the human ear is most sensitive to frequencies in the higher range of audible frequencies, tosubjective assessment of noisiness it is necessary to reduce the sound pressure levels of the lowfrequency bands before combining them with the higher octave frequency bands. Corrections for thebands of interest for the A, B and C weighting curves are given in Table 9.1. In using these curves therelative sound level is added (arithmetically) to the band sound pressure level to give the weightelevel. The weighted band levels are then summed to give the weighted sound pressure leveAppendixA or Equation (9.3).

, (9.3)

i.e. , (9.4)

where is the weighted sound pressure level and are the weighted band sound pressulevels in each of the octave bands. After summation of the band levels, the resulting noise level isas dBA, dBB and dBC respectively for the A, B and C curves. The most commonly used sound plevels are A-weighted.

As an alternative to weighted sound pressure levels, the annoyance of fan system noise may be using noise criteria or noise rating curves. Noise assessment on the basis of these curves is more inby the spectral content of the noise than a weighted sound pressure level. Details of noise criteria arating curves are given in Derivation 10 and Reference 5 respectively.

9.4 Summation of Sound Power Levels

The total sound power level from more than one source may be found by first converting the indisound power levels to sound power, then summing* (arithmetically) to obtain the total sound power (

* The sound power from contrarotating axial fans may not be obtained in this way because the closely spaced impellers interfere. On averagea two-stage axial fan with contrarotating blades will be 5 to 6 dB noisier than a guide-vane fan working at the same duty.

Lower band limit, Hz 44 88 177 355 710 1420 2840 5680

Centre frequency, Hz 63 125 250 500 1000 2000 4000 8000

Upper band limit, Hz 88 177 355 710 1420 2840 5680 11 360

fcentre fupper flower×( )½=

Lp

Lp 10 antilog10

Lp1

10--------- antilog10

Lp2

10--------- ... antilog10

Lp8

10---------+ + +log10 =

Lp 10 10Lpj 10⁄

j=1

8

∑log10 =

Lp Lp1 L, p2 ... Lp8,

59

79037�cally as

taveay be

ethods,s ofrivationd centred at 63rovided pressured using

ing uponned, andwith the uniform

radiated to the

sound

l is given

empstsound

watts) and finally reconverting to a total sound power level. This procedure is expressed mathemati

total sound power level

, (9.5)

i.e. total sound power level , (9.6)

where is the number of sources and is the sound power level of the source. A graphical procedureis given for this summation in Appendix A. The summation should be carried out separately for each ocband; however, if a single overall sound power level is required the octave band power levels msummed using the same procedure.

9.5 Fan Noise Testing

The three preferred methods for measurement of fan noise, in-duct, free-field and reverberant-field mare fully described in Derivation 16. The primary objective of fan noise testing is to ascertain valuesound power level in each of the octave frequency bands of interest. It should be noted that in De16 fan manufacturers are only required to provide noise data for the octave bands in the range of banfrequencies from 125 Hz to 4000 Hz. Many manufacturers also provide data for the bands centreHz and 8000 Hz but, due to difficulties in measuring noise levels in these latter bands, the data pmay not be guaranteed to the same accuracy as those given for the other octave bands. Soundlevels are first measured in each octave band and the equivalent sound power level estimateappropriate relationships*.

Measurements to determine either the inlet or outlet fan sound power levels may be made. Dependthe test method used, either inlet sound power levels or open-inlet sound power levels are determisimilarly for outlet and open-outlet sound power levels. These sound power levels are associated appropriate sound powers. The inlet sound power of a fan is the rate at which sound energy enters aairway connected to the fan inlet; the open-inlet fan sound power is the rate at which sound energy isinto the ambient atmosphere from the open-inlet and the casing of the fan. Similar definitions applyoutlet and open-outlet sound powers.

A means of conversion between inlet (or outlet) sound power levels and open-inlet (open-outlet)power levels, for ducts of uniform cross section terminating in open space, is provided in Figure 5.

9.5.1 In-duct fan testing

For in-duct testing the relationship between measured sound pressure level and sound power leveby

, (9.7)

where is the cross-sectional area† of the test airway.

* All equations given in this Item relating sound pressure level to sound power level are approximate since these relationships are dependenton the temperature and pressure at the time of measurement. For general engineering applications within the usual range of teratureand pressure, this dependence can be ignored. Further, it is assumed that sound pressure level measurements are taken at a diance remotefrom the sound source (i.e. acoustic far-field) such that particle velocities are predominantly in the direction of propagation of the and the acoustic intensity is proportional to sound pressure squared. It is usual to assume that far-field conditions exist at distances morethan one sound wavelength away from the source or at 3 times the fan diameter, whichever is the greater.

10 antilog10

Lw1

10--------- antilog10

Lw2

10--------- .... antilog10

LwJ

10---------++log10 =

10 10Lwj 10⁄

j=1

J

∑log10 =

J Lwj j th

Lw Lp 10 Alog10 +=

A

60

79037�

tions or soundnd for thend source.rithmic

and,

nships:

.

diffuse, is

easured

sing the

e in anre in the

els forin eachs soundeasured.entative

ed.

9.5.2 Free-field fan testing

For free-field testing the fan assembly may be mounted such that sound radiates freely in all direcit may be mounted on a flat sound reflecting surface within a free-field test facility. For the first casepressure levels are measured at points on a hypothetical sphere centred on the sound source alatter case the sound measuring points are located on a hypothetical hemisphere centred on the souFor either testing method, the sound pressure level within an octave frequency band, , is the logamean of the values ( ) measured on the sphere, or hemisphere, for that frequency bi.e.

. (9.8)

The open-inlet or open-outlet sound power level is then estimated using one of the following relatio

for spherical sound radiation

, (9.9)

and for hemispherical sound radiation

, (9.10)

where for both expressions is the radial distance of the measuring points from the sound source

9.5.3 Reverberant-field fan testing

For reverberant-field testing, sound pressure levels are measured at a number of locations in thesound field. In each octave frequency band the mean reverberant-field sound pressure level, determined from the averaged value of the several band sound pressure levels ( ,) mwithin the test chamber, using Equation (9.8) where is the number of measurement locations.

The open-inlet or open-outlet sound power levels in each octave band are then estimated urelationship

, (9.11)

where is the volume of the test enclosure and is the reverberation time. The reverberation timenclosure, for sound of a given frequency, is the time required for the mean square sound pressuenclosure to decay 60 dB from its initial steady state after the source is stopped.

9.6 Understanding Fan Noise Data

In providing noise data for their fans, manufacturers usually quote sound power (or pressure) levvarious operating conditions. They also provide a spectrum shape so that the sound levels withoctave band may be deduced from the overall levels. Particularly when noise data are provided apressure levels, attention should be given to the test conditions under which those data were mSound pressure levels measured in free-field or reverberant-field conditions are not generally repres

† In using the relationships in this Item for estimating values of from measured values of , the SI system of units is assumWithalternative unit systems these relationships have additional constant terms.

Lw Lp

Lpmj Lp1 Lp2, ... Lpj,

Lpm 101j--- antilog10

Lp1

10--------- antilog10

Lp2

10--------- ... antilog10

Lpj

10-------+ + +log10 =

Lw Lpm 20 r 11+log10 +=

Lw Lpm 20 r 8.0+log10 +=

r

LpmLp1 Lp2, ... Lpj,

j

Lw Lpm 10 T 10 V 14–log10 +log10 –=

V T

61

79037� Soundrried out.

ced to aditions,n sound

power sound equally

d have from one

inationen inlet

n Figuren due to

dure itst whenrticularements,ocedureobtaind in the

ce thisto the fan

er levels,nciesensitive flowquency

be less

onse toring bydiction,

ed here.d powery.

of those achieved in practice since conditions in most buildings fall between those two extremes.pressure levels must be converted to sound power levels before system noise calculations can be ca

Noise data given in terms of sound pressure levels for in-duct or free-field noise tests are referenparticular distance from the sound source (usually three fan diameters or metres). For free-field consound pressure levels may be corrected to any required reference distance since for that conditiopressure level is inversely proportional to the square of the distance to the noise source.

Fan manufacturers generally quote either the fan inlet or fan outlet sound power levels. If total soundlevel is quoted, it is the arithmetic average of the total inlet sound power level and the total outletpower level. In the absence of further data it is usual to assume that fan sound power is emittedfrom the fan inlet and outlet.

When comparing the noise specifications for two fans, it is important to ensure that the data providea common basis and have been measured under similar conditions. For example, outlet noise datafan should not be compared with open-outlet data from another, since end reflections in the duct termon the former fan will attenuate the sound radiated into open space. A means of conversion betwe(or outlet) and open-inlet (or open-outlet) sound pressure levels, based on duct area, is provided i5 for the range of octave frequency bands of interest. This figure shows that the greater attenuatioduct end reflections occurs at the lower frequencies.

In using this Item to estimate fan sound power levels it should be noted that in the prediction proceis assumed that the fan is working at, or close to, its maximum efficiency. Fans are at their quieteoperating near peak efficiency and noisiest when working at, or near, their stalled condition. For a painstallation, it is therefore necessary to select the type of fan on the basis of the performance requirdiscussed elsewhere in this Item, before considering noise criteria. Because the prediction prprovided considers the fan to be working at its maximum efficiency it is better, where possible, to fan noise data from manufacturers' catalogues for the fan operating at the specific duties requireparticular installation.

It should be noted that fan generated noise will increase if the airflow into the fan is unsteady. Sinincrease may be considerable, to ensure its quietest operation it is essential that the air passing inis uniform and turbulence free.

The noise generated by the basic types of fan may be compared by considering the specific fan pow, given in Table 9.2. It is evident that centrifugal fans generate most of their noise at low freque

while axial-flow fans generate high noise levels at high frequencies. Since the human ear is more sto high frequency noise, axial fans will be noisier than centrifugal fans operating under similarconditions. However, it should be noted that it is easier to suppress high frequency noise than low frenoise. It therefore follows that a high-speed axial-flow fan fitted with an appropriate attenuator may noisy, and cheaper to install, than a slower speed centrifugal fan giving the same performance.

When selecting a fan on the basis of noise it is important to realise that, although the human respnoise is frequency dependent, it is not, in general, possible to distinguish between noise levels diffeless than about 3 dB. Also because of the inherent difficulties in both fan noise measurement and predifferences in noise levels of less than 2 dB are usually considered insignificant.

9.7 Estimation of Fan Noise

The in-duct sound power levels of fans within ducts may be estimated using the procedure describFor the case of propeller fans the predicted sound powers are the open-inlet (or open-outlet) sounlevels. It is implicit in this prediction procedure that the fan is working at, or near, its peak efficienc

Kw

62

79037� system sound

ande bands

from 63

erating

nt of the

passing

y band

sidered

m haveues forverallividual centred

The in-duct sound power levels include, at the lower frequencies, the sound energy within the ductwhich does not emerge into the occupied space owing to reflections at the outlet. The reduction inpower levels at the outlet (i.e. the difference between open-inlet (or open-outlet) sound power levelinlet (or outlet) sound power level) is plotted against duct cross-sectional area for each of the octavin Figure 5.

The sound power level of a fan in each octave frequency band, in the range of centre frequencies Hz to 8000 Hz, is given by

. (9.12)

The base sound power level, , which is independent of frequency, is found from the fan opcondition. Values of are plotted against for a range of values of in Figure 7. The specific fansound power level, , accounts for spectral variations in the sound power level and is independefan operating conditions. Values of for each octave frequency band are given in Table 9.2 for the varioustypes of fan.

It is necessary to apply a correction to the sound power level in the octave band containing the bladefrequency, . The blade passing frequency is given by

. (9.13)

The octave band containing may be obtained from Figure 6.

Having identified the octave frequency band containing , the sound power level in that frequencis augmented by the blade frequency increment given in Table 9.2.

For guidance in the initial system design, typical numbers of fan blades for the various fan types conare given in Table 9.2.

Overall sound power levels and octave band sound power levels predicted by the method in this Itebeen compared with the equivalent in-duct sound power levels given in fan manufacturers’ catalogfans working at their peak efficiency. On the basis of that comparison, the majority of predicted osound power levels are within dB of the catalogue levels. The scatter of compared data in the indoctave bands is, in most cases, greater than that for the overall level; for example, in the octave bandat 250 Hz, the majority of predicted levels are within dB of the catalogue levels.

TABLE 9.1

OctavecentrefrequencyHz

A-weighting relative sound pressure leveldB

B-weighting relative sound pressure leveldB

C-weighting relative sound pressure leveldB

63

125

250 0

Lw Kw Lw∗+=

Lw∗

Lw∗ q ∆pf t,

KwKw

fB

fBfan speed (rev/min) number of blades×

60------------------------------------------------------------------------------------------------=

fB

fB

7±

9±

26.2 – 9.3– 0.8–

16.1– 4.2– 0.2–

8.6– 1.3–

63

79037�

blades

* When a range is given for the blade frequency increment, the higher increments occur for fans having the lower number ofand the lower increments for fans having the higher number of blades.

500 0

1000 0 0 0

2000

4000

8000

TABLE 9.2

Fan type

dB Blade* frequency increment

dB

Typical number of fan blades

Octave band centre frequencies Hz

63 125 250 500 1000 2000 4000 8000

Propeller 51 48 49 47 45 45 43 31 5 – 7 3 – 7

Axial Tube axial 44 44 44 44 41 38 32 25 6 – 8

7 – 16Guide-vane axial

43 42 42 43 41 38 34 26 6 – 8

Forward-curved

43 41 36 32 28 24 20 15 2 32 – 64

Centrifugal Radialblade

46 47 47 43 39 32 25 19 2 – 5 5 – 10

Backward-bladed

39 36 34 32 28 24 18 12 3 8 – 16

Mixed flow Axialcasing

46 43 43 38 37 32 28 25 4 – 6 8 – 15

TABLE 9.1

3.2– 0.3–

1.2+ 0.1– 0.2–

1.0+ 0.7– 0.8–

1.1– 2.9– 3.0–

Kw

64

79037�

d flow

for fans is

hus,o that

e

-

must

ad as

10. EXAMPLES

10.1 Fan Selection and Size and Power Estimation

A fan is required to supply air to a small, fixed, ventilation system installed at sea level. The requirerate is 8 m3/s for which a gauge static pressure of 500 Pa is needed.

Sketch 10.1 Gauge pressure, = 500 Pa required

Select a suitable fan, obtain its approximate dimensions and estimate the power required.

Within Section 7 on fan selection, Section 7.1 gives guidance on the mechanical suitability of fans different duties. Since the conditions in this example are not mechanically arduous, the choice ofnot restricted by such considerations.

Section 7.2 explains the principles for sizing the various types of fans and Section 7.2.1 details the procedureas follows.

.

Step Procedure

1. The notes in Section 4 and in particular Section 4.2.1 state that forward-curved centrifugal fans arewell suited to applications requiring a high volume flow rate at low to medium pressure rises. Tthis type of fan appears appropriate especially as it draws in air directly from the atmosphere sno special inlet connections have to be made.

2. A middle-range value from Figure 1 of the parameter for forward-curved fans is 1.2. Thcorresponding value of .

3. The notes on Figure 1 advise that a value for of approximately 0.5 is typical for forwardcurved fans.

4. The relative density, , has a value of unity since the fan is to be installed at sea-level.

5.–7. Following the dashed line representing this example in Figure 1, a value for of 500 and m3/s locates point A on the upper right-hand graph. Note that the fan static pressure

be equal to 500 Pa for the reasons set out in Section 3.1.2 and, in particular, Sketch 3.3a.

6.. Point B is located on the Reference Line for centrifugal fans.

7. Point C is located on the central inclined grid and the corresponding value for may be re m.

8. Point A is pinpointed on Figure 2.

ps

1 ds⁄ns 2.7=

WT DT⁄

σ

∆pf s , σ⁄qmax 8 =

DTDT 0.74=

65

79037�

s will be

ns and

total

er leveldes.

.

ead as

nd

read

f the

As the impeller speed and diameter appear reasonable at this stage, the fan external dimensionestimated. Section 7.2.2 tables approximate dimensions of the fan casing. From Section 7.2.2 and usingtypical middle-range values, the estimated fan dimensions are as follows.

Section 7.3 explains the principles for estimating the power requirements for the various types of faSection 7.3.1 details the procedure as follows.

The fan total pressure rise may also be estimated using the correlation given in Figure 4. For this examplewith , the value of for a forward-curved fan is approximately 0.7. Thus the fan pressure rise is given by Pa.

10.2 Fan and System Noise Estimation

10.2.1 Fan noise estimation

For the fan of the previous example running at the same conditions, estimate the in-duct sound powin each of the eight octave frequency bands and the overall sound power level. The fan has 40 bla

9. Using = 2.7 as before, point B is located on the Reference Line for the rev/min speed scale

10. Point C is located on the central inclined grid and the corresponding impeller speed may be r606 rev/min.

11. The fan will be belt driven and so will not be restricted to electric motor pole speeds.

The overall height, , is m,

the overall depth, , is m,

and the casing width, , is m.

Step Procedure

1. From Figure 3, for a value of , the fan static efficiency may be read from the far left-hagraph as 42 per cent.

2. Point A is located on the far right-hand graph corresponding to Pa and m3/s.

3. Point B is located on the Reference Line for centrifugal fans.

4. Point C is located within the central inclined grid and the corresponding value for may beas W.

5. A higher efficiency could have been obtained from the same type of fan (forward curved) ispecific speed were reduced. However, as an inspection of Figure 1 reveals, this would have resultedin a larger fan.

Step Procedure

ns

A 1.95 0.74× 1.4=

B 1.7 0.74× 1.2=

C 1.2 0.5× 0.5+( )0.74 0.8=

ns 2.7=

∆pf s, 500= qmax 8 =

PRPR 9500 =

ns 2.7= ∆pf s , ∆pf t,⁄∆pf t, 500 ≈ 0.7⁄ 715 =

66

79037�

uency

ncy.

nd

given in

alue given

Using Figure 7 with m3/s and Pa.

.

From Figure 6, for a fan with 40 blades rotating at a speed of 606 rev/min, the blade passing freqoccurs in the octave band centred at 500 Hz.

For a centrifugal fan having forward-curved blades, the blade frequency increment, given in Table9.2, is2 dB.

The sound power level for each octave band is found using the relationship

and applying the appropriate correction to the frequency band containing the blade passing freque

Values of for the particular fan type are found from Table 9.2 and the summation for each octave bais tabulated.

To obtain the overall sound power level, the eight octave band levels are combined by the method Appendix A or by using Equation (9.6).

Using the method in Appendix A successive pairs of sound power levels are combined until a single vis obtained. The combination of levels is tabulated here and the detail of part of this combination isin the example in Appendix A.

Octave bandcentre frequency

Hz 63 125 250 500 1000 2000 4000 8000

dB 51.5 51.5 51.5 51.5 51.5 51.5 51.5 51.5

Blade frequencyincrement dB 2

dB 43 41 36 32 28 24 20 15

In-duct band soundpower level dB 94.5 92.5 87.5 85.5 79.5 75.5 71.5 66.5

q 8 = ∆pf t, 715 =

Lw∗ 51.5 =

Lw Lw∗ Kw+=

Kw

Lw∗

Kw

67

79037�

nal areaectionst outlet.ttern is

or losses

herical

e sound

The estimated overall sound power level in the duct is 97.5 dB.

10.2.2 System noise

The fan of the previous example services a space through a length of uniform duct of cross-sectio0.02 m2 which terminates in the centre of a wall. Neglecting duct losses, other than those due to reflat the duct termination, estimate the sound pressure level in dBA at a distance of 3 m from the ducAssume that free-field conditions exist in the serviced area and that the sound radiation panon-directional.

Firstly the sound power radiated into the space is estimated by correcting each octave band level fdue to end reflections at the duct exit. These losses are obtained from Figure 5 for the duct exit area of 0.02m2.

Since the duct exit is in the centre of a wall, and the radiation pattern is non-directional, hemispradiation can be assumed and the octave band sound pressure levels at a distance of 3 m (i.e. ) maybe found using Equation (9.10),

,

i.e. . (10.1)

The sound pressure levels in each octave band are then weighted using the A-weighting relativpressure levels given in Table 9.1.

Finally the weighted band levels are summed, using the method in Appendix A or Equation (9.4), to obtainthe single-value sound pressure level in dBA.

The calculation steps are tabulated.

r 3=

Lw Lpm 20 r 8.0+log10 +=

Lpm Lw 20 r 8.0–log10 –=

68

79037�

vel, the

The end correction for the 63 Hz centred band falls outside the range of Figure 5. Since the correction inthis frequency band will reduce the band sound pressure level well below the maximum band lecontribution of the 63 Hz centred band to the dBA sound pressure level is negligible.

The summed A-weighted sound pressure level is 65.6 dBA.

Octave band centrefrequency

Hz 63 125 250 500 1000 2000 4000 8000

In-duct sound power levelfrom previous Example dB 94.5 92.5 87.5 85.5 79.5 75.5 71.5 66.5

End correction from Figure5

dB see below

14.7 9.2 4.6 1.2 0 0 0

Sound power level radiatedinto serviced area dB – 77.8 78.3 80.9 78.3 75.5 71.5 66.5

Sound pressure level at 3 mfrom exit. Equation (10.1) dB – 60.3 60.8 63.4 60.8 58.0 54.0 49.0

A-weighted relative soundpressure level from Table9.1

dB –26.2 –16.1 –8.6 –3.2 0 +1.2 +1.0 –1.1

A-weighted band soundpressure level dB – 44.2 52.2 60.2 60.8 59.2 55.0 47.9

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11. REFERENCES AND DERIVATION

11.1 References

The references given are recommended sources of information supplementary to that in this Item (chronological order).

11.2 Derivation

The derivations given are sources of information, excluding manufacturers' catalogue data, emplthe production of this Item (listed in chronological order).

1. – Classification of insulating materials for electrical machinery aapparatus on the basis of thermal stability in service. British StandInstitution, B.S.2757, 1956.

2. – Method of rating industrial noise affecting residential and industareas. British Standards Institution, B.S.4142, 1967.

3. SHARLAND, I. Woods practical guide to noise control. Woods of Colchester L1972.

4. – Motor starters for voltages up to and including 1000 V a.c. and 120d.c. Part 1 - Direct online (full voltage) a.c. starters. British StandaInstitution, B.S.4941, Pt 1, 1973.

5. – Handbook of noise and vibration control. Edited by R.H. Warring.Trade and Technical Press Ltd, 1973.

6. IQBAL, M.A.WILLSON, T.K.THOMAS, R.J.

The control of noise in ventilation systems - a designer's guide. E. andF.N. Spon, London, 1977.

7. OSBORNE, W.C. The selection and use of fans. Engineering Design Guides, NDesign Council, 1979.

8. ESDU Fluid mechanics, Internal Flow sub-series, Vols 1-4, EngineeSciences Data Unit, 1980.

9. – Motor starters and controllers. British Standards Institution, B.S.51957

10. – Noise reduction. Edited by L.L. Beranek. McGraw-Hill, 1960.

11. – Woods practical guide to fan engineering. Edited by W.C. Osboand C.G. Turner. 2nd Edition, Woods of Colchester, 1960.

12. – Methods of testing fans. Part 1 – Performance. British StandInstitution, B.S.848, Pt 1, 1963.

13. HUGHES, E. Electrical technology. 4th Edition, Longman, 1969.

14. – Fan engineering. Edited by R. Jorgensen. Seventh Edition, BuffaForce Company, USA, 1970.

70

79037�

ards

ow

ers

st),

tics.

ust

9.

15. CSANADY, G.T. Theory of turbomachines. McGraw-Hill Book Co., USA, 1964.

16. – Method of testing fans. Part 2 - Fan noise testing. British StandInstitution, B.S.848, Pt 2, 1966.

17. – Design for sound. Woods Fans, Toronto, April 1966.

18. KENNY, R.J. Fans and blowers. Machine Design, Vol.40, No.6, pp.l52-173, March1968.

19. BUSH, E.H. Crossflow fan - history and recent developments. Fan Technology andPractice, Paper 4, pp.50-66, Inst. mech. Engrs, April 1972.

20. GRAHAM, J.B. How to estimate fan noise. Sound and Vibration, Vol.6, No.5,pp.24-27, May 1972.

21. ECK, B. Fans - design and operation of centrifugal, axial-flow and cross-flfans. 1st English edition, Pergamon Press, 1973.

22. FINKELSTEIN, W. Preliminary computations of fan noise. Bldg Services Engr, Vol.41,pp.268-275, March 1974.

23. SEARLE, D.G. Pole amplitude modulated motors for fan and pump drives. ElectricalReview, Vol.194, pp.407-409, 19th April 1974.

24. – Fan application guide. Heat. Vent. Air Conditioning ManufacturAssn, 1975.

25. – Fan terminology. Eurovent 1/1, Paris, March 1975.

26. – Electrical motors and controls, Section 1 Motors. Machine Design,Vol.48, No.10, pp.8-42, 1976.

27. ASTROM, L. Fans for VAV systems. Flakt Review, Vol.13, No.2, pp.18-22, 1978.

28. GORDON, C.G. Fan noise and its prediction. Internoise 78, pp.154-165, SanFrancisco, May 1978.

29. PAPAMARCOS, J.BAESEL, H.D. et al.

Choosing fans for power plants. Power Engineering, Vol.82, No.6,pp.46-60, June 1978.

30. GERRARD, G.W.MESSER, M.G.M.

Designing with moving air. Original Equipment Manufacture Design,Parts 1 to 5, pp.24-26 (June), pp.36-41 (July), pp.37-41 (Augupp.50-55 (September), pp.53-59 (October), 1978.

31. McFARLAND, J. Fans for filtration. Filtration and Separation, Vol.16, No.2,pp.l44-149, March/April 1979.

32. – Specification for quantities, units and symbols. Part 7 - AcousBritish Standards Institution, B.S.5775, Pt 7, 1979.

33. BUSH, E.H. Private Communication, Airwheel Ltd, Holton Heath, Dorset, Aug1979.

34. CAMPBELL, J. Private Communication, Ove Arup Partnership, November 1979.

35. WOODS-BALLARD, W. Private Communication, Woods of Colchester Ltd, December 197

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11.3 Manufacturers' Catalogue Data

The following manufacturers, listed in alphabetical order, supplied catalogue and other data emplthe production of this Item.

C1. AIRSCREW Mixflo fans. Airscrew Howden Ltd, Weybridge, Surrey, UK.

C2. AIRWHEEL L, S and U flow fans. Leaflets on tangential fans. Airwheel LHolton Heath, Dorset, UK.

C3. BLACKMAN Series KB 24 centrifugal fans. Publication 65, 2nd Edition. KeBlackman Ltd, London N17, UK.

C4. DAVIDSONS Standard fan catalogues on Sirocco axial fans and Sirocco centrifans. Davidson and Co. Ltd, Belfast, Northern Ireland.

C5. ENGART Technical literature for HD and HV series of centrifugal fans. EngFans Ltd, Aberdare, UK.

C6. FLAKT FTBA axial-flow fans and HCL, HCM and HCH centrifugal fanFlakt Ltd, Staines, Middlesex, UK.

C7. HALIFAX Paddle, backward inclined, backward curved and multivane faHalifax Fan Manufacturing Co. Ltd, Salterhebble, Halifax, UK.

C8. ITT Medium pressure tangential blowers. ITT, Thornton Industrial EstMilford Haven, Dyfed, UK.

C9. JOY Joy axivane fans. Catalog J-610, Joy Manufacturing Co., Ohio, US

C10. MATTHEWS & YATES Centrifugal fans. Catalogue F/7, Matthews and Yates, SwinManchester, UK.

C11. MYSON Axial-flow fans and propeller fans. Myson Fans Ltd, Colchester, U

C12. SMITHS Propeller fan PFD catalogue and leaflets on cross-flow fans. SmIndustries Precision Fan Co., Witney, Oxon, UK.

C13. STANDARD ANDPOCHIN

Centrifugal fan catalogue. Standard and Pochin Ltd, Leicester, UK

C14. TROX Axial fan data sheets. Trox Bros Ltd, Thetford, Norfolk, UK.

C15. WOODS Aerofoil axial fans: performance charts and data tables, PublicAFI, Woods of Colchester Ltd, UK.

72

79037�

4 610

2 4 6102 2 4 6

103 2 4 6

3/s)

10002000500010000

se of 7.2.1

e wne

al w

!

!

!

73

FIGURE 1 ESTIMATION OF IMPELLER TIP DIAMETER, DT

10-3 2 4 610-2 2 4 6

10-1 2 4 61

2

qmax (m

0.6 0.8 2 4 6 8 20 401 10 10-2 2 4 610-1 2 4 6 1

2 4 6 10 20ns DT(m)

0.1

0.2

0.4

0.6

0.8

2

4

6

8

1

10

WT / DT

50100200500

∆pf,s /σ (Pa)

For guidance on the uthis chart, see Section

ta

mfpa

va

pa

va

ta

mfcr

0.2

0.4

0.6

0.8

1

2

Section

4.1.14.1.24.1.34.1.4

4.2.14.2.24.2.24.2.34.2.34.2.3

4.3.1

4.4.1

Legend

pa propeller fanta tube-axialcr contrarotatingva guide-vane axial

fc forward-curved centrifugalpb radial-discharge, paddle bladedch radial-discharge, curved heelba backward-bladed, aerofoilbc backward-bladed, curvedbi backward-bladed, straight

mf mixed-flow, axial casing

xf cross-flow, J,S and U types

Note:

For all axial and mixed-flow fans,take WT / DT = 1

For single-inlet centrifugal fans, typicalranges of WT / DT are as follows:

forward-curved 0.40 < WT / DT < 0.65

radial discharge 0.40 < WT / DT < 0.55

backward bladed 0.12 < WT / DT < 0.30

For double-inlet, double the above values.For cross-flow fans, 0.70 < WT / DT < 10

increasing dimeterincreasing speed

peak efficiency fans availableoff-peak efficiency but innormal operating range

0.4

0.6

0.8

1.0

2

1 / ds

1 / ds

Re

fere

nce

lin

for

axi

al

an

d m

ixe

d-f

lofa

ns

Re

fere

nce

lifo

r ce

ntr

ifu

ga

nd

cro

ss-f

lofa

ns

!

!

!

!

!

!

fc

chba, bc, bi

ba, bc, bi

xf

xf

pb

pb

74

79037�

10 102 103 1046 8 2 4 6 2 4 6 2 4 6

A

100005000

00

∆pf,s /σ (Pa)

For guidance on the use ofthis chart, see Section 7.2.1

!

!

!

/s)

FIGURE 2 ESTIMATION OF IMPELLER SPEED, N

10-3 10-2 10-1 12 42 4 6 82 4 62 4 6

0.6 0.61 2 4 6 810 20 4010

2 4 6102 2 4 6

103 2 4 6 104 2

(Note value of ns must

be the same as that used in Figures 1,3 and 4 )

ns Impeller speed rad/s or rev/min

Note:

For all axial and mixed-flow fans,take WT / DT =1

For single-inlet centrifugal fans, typicalranges of WT / DT are as follows :

forward-curved 0.40 < WT / DT < 0.65

radial discharge 0.40 < WT / DT < 0.55

backward bladed 0.12 < WT / DT < 0.30

For double-inlet, double the above values.For cross-flow fans, 0.70 < WT / DT < 10

0.1

0.2

0.4

0.60.81

2

4

6810

WT / DT

B

C

201000

500200

10050

Re

fere

nce

lin

efo

r re

v/m

inim

pe

ller

spe

ed

Re

fere

nce

lin

efo

r ra

d/s

imp

elle

r sp

ee

d

!

!

!

!

!

qmax (m3

79037�

ce on the use ofsee Section 7.2.1

1000

050

0020

0010

00

200

100

50

500

∆pf,s (Pa)

qmax (m3/s)

!

Legend Section

A

1 10 102 103642 642 642642

75

FIGURE 3 ESTIMATION OF IMPELLER POWER, PR

For guidanthis chart,

ηs

per cent

ηsper cent

100

10080

60

40

20

80

60

40

20

PR (watts)

!

!

!

!

!

ta

pavacr

mf

mfpa

crvata

xf

pb

pb

fc

chbi

bc

ba

pa propeller fanta tube - axial cr contra - rotatingva guide - vane axial

fc forward - curved centrifugalpb radial - discharge, paddle bladedch radial - discharge, curved heelba backward - bladed, aerofoilbc backward - bladed, curvedbi backward - bladed, straight

mf mixed - flow, axial casing

xf cross - flow, J,S and U types

4.1.14.1.24.1.34.1.4

4.2.14.2.24.2.24.2.34.2.34.2.3

4.3.1

4.4.1

ns

(Note value of ns must be the same

as that used in Figures 1, 2 and 4)

Re

fere

nce

lin

efo

r ce

ntr

ifug

al a

nd

cro

ss -

flo

w f

an

s

Re

fere

nce

lin

efo

r ax

ial a

nd

mix

ed

- f

low

fa

ns

C

B

1 10 10 102 103 104 105 106 10-3 10-2 10-10.5 2 4 6 20 40 2 4 6 2 4 6 2 4 6 2 642642642642

Peak efficiency fans availableOff - peak efficiency, but in normal operating range

76

79037�

25 30 40 50

of ns must be the

at used in Figures 1,2 and 3 )

eak efficiency fans available

ff-peak efficiency but inormal operating range

Legend Section

propeller fantube - axialcontrarotatingguide-vane axial

forward-curved centrifugalradial-discharge, paddle bladedradial-discharge, curved heelbackward-bladed, aerofoilbackward-bladed, curvedbackward-bladed, straight

mixed-flow, axial casing

ross-flow, J,S and U types

4.1.14.1.24.1.34.1.4

4.2.14.2.24.2.24.2.34.2.34.2.3

4.3.1

4.4.1

FIGURE 4 CORRELATION OF

ns

0.5 0.6 0.7 0.8 0.9 1.5 2 2.5 3 4 5 6 7 8 9 15 201 10

∆pf,s / ∆pf,t

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1.0pb

xf fc

cr mf

ba,bc,bi

va

ta

ba,bc,bi

mf

va

cr

ch

ta

pbxf

fc

( Note value

same as th

p

on

pa

pa

ch

pa ta cr va

fc pb ch ba bc bi

mf

xf c

∆pf s,∆pf t,--------------

79037�

FIGURE 5

Duct area m2

2 3 4 6 8 2 3 4 6 8 2 3 4 6 8 2 3 4 510-2 10-1 100

Inle

t (or

ou

tlet)

sou

nd p

ower

leve

l -op

en in

let (

or o

pen

ou

tlet)

sou

nd p

ower

leve

l. d

B

0

2

4

6

8

10

12

14

16

18

20

63

125

250

500

1000

20004000

Octave bandcentre frequency Hz

77

79037�

FIGURE 6

6

8

102

103

2

3

4

6

8

2

Number of blades

605040

30

20

15

10

7

54

3

2

Fan speed rev/min102 103

fB Hz

Octavebandsandcentrefreq. Hz

2000

1000

500

250

125

63

6 8 2 3 4 6 8 2 3 4

78

79037�

4 6 8 2 3 4 6 8103 104

100

104

103

102

2

2

4

4

6

6

8

8

∆pf,t Pa

79

FIGURE 7

q m3/s

2 3 4 6 8 2 3 4 6 8 2 3 4 6 8 2 3 4 6 8 2 310-2 10-1 100 101 102

Lw*

dB

0

10

20

30

40

50

60

70

80

90

79037�

vels isny other

eference

rces

vels inver.

vels

levels

APPENDIX A COMBINATION OF LEVELS IN dB *

A1. NOTATION

A2. NOTES

A method of determining the combined sound level resulting from two or more sources of known lepresented in this Appendix. The method is applicable to sound power levels and pressure levels or aparameter expressed in dB provided that all levels to be combined are with respect to a common rvalue.

In the Figure A1, is plotted against . Clearly the combined level from any number of soumay be obtained by repeated combination of pairs of levels.

The method may be used to obtain the overall level for a single source by combination of the leparticular frequency bands. The method applies strictly to statistically independent sources, howe

Figure A1 shows that is negligible when is large, rising to a maximum of 3 dB when the leare the same.

A3. DERIVATION

.

A4. EXAMPLE

It is required to determine the overall sound power level resulting from three sources with individualof 94.5, 92.5 and 89.6 dB with respect to a common reference power.

To combine the first pair of levels, choose

dB, dB,

so dB.

* This Appendix is also available as Item No. 66017 in the Noise Sub-series and in the Acoustic Fatigue Sub-series.

higher level dB

lower level dB

level due to combination of and , dB

increase in level above due to dB

1. – Handbook of noise control. Edited C.M. Harris. McGraw-Hill, NewYork, 1957.

L1

L2

L L1 L2 L1 L1∆+

L1∆ L1 L2

L1∆ L1 L2–

L1∆ L1 L2–

L1∆ 10 1 antilog10

L2 L1–( )

10-------------------------+

log10 =

L1 94.5= L2 92.5=

L1 L2– 2.0=

80

79037�

From Figure A1 dB.

Therefore dB.

Now combining this with the remaining original level, choose

dB, dB,

so dB.

From Figure A1 dB.

Therefore dB,

which is the overall level for the original three sources.

L1∆ 2.1=

L 94.5 2.1+ 96.6= =

L1 96.6= L2 89.6=

L1 L2– 7.0=

L1∆ 0.8=

L 96.6 0.8+ 97.4= =

81

82

79037�

16 18 20

FIGURE A1 COMBINATION OF LEVELS IN dB

L1 - L2 dB

0 2 4 6 8 10 12 14

∆ L1 dB

0.0

0.5

1.0

1.5

2.0

2.5

3.0

79037�

ar Item

eeringhanicalpointed

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eeringinitialtem were

THE PREPARATION OF THIS DATA ITEM

The work of the permanent professional staff of the Engineering Sciences Data Unit on this particulwas monitored and guided by the following Working Party.

Mr A.R. Green – Trox Brothers LtdProf. R.I. Lewis – Newcastle UniversityMr J. McFarland – Matthew & Yates LtdMr E.J. Perry – Atkins Research and DevelopmentDr D. Pollard – GEC Power EngineeringMr W.R. Woods-Ballard– Woods of Colchester Ltd.

on behalf of the Internal Flow Panel which has the following constitution:

ChairmanMr N.G. Worley – Babcock Power Ltd

MembersMr J. Campbell – Ove Arup PartnershipDr D. Chisholm – National Engineering LaboratoryDr D.J. Cockrell – Leicester UniversityDr R.B. Dean – Atkins Research and DevelopmentMr D.H. Freeston* – Auckland University, New ZealandDr G. Hobson – GEC Turbine Generators Ltd. RugbyProf. J.L. Livesey – Salford UniversityMr D.S. Miller – British Hydromechanics Research AssociationMr B. Payne – Kellogg International CorporationDr D. Pollard – GEC Power Engineering Ltd. WhetstoneMr J.A. Ward – Atomic Energy Technology Unit.

The Internal Flow Panel has benefited from the participation of members from several engindisciplines. In particular, Dr G. Hobson has been appointed to represent the interests of mecengineering as the nominee of the Institution of Mechanical Engineers and Mr B. Payne has been apto represent the interests of chemical engineering as the nominee of the Institution of Chemical En

The work on this Item was carried out in the Internal Flow and Physical Properties Group of the EnginSciences Data Unit. The members of staff who undertook the technical work involved in the assessment of the available information and the construction and subsequent development of the I

Mr C.J.T. Clarke – Group Head, Internal Flow and Physical PropertiesMr R.F. Lambert – Group Head, Noise and Structural DynamicsMr M. Bolton – Senior Engineer, Internal Flow and Physical Properties Group.

* Corresponding Member

83