tlt english

1 Elementary Fan Technology

2

Prof. Dr.-Ing.Reinhard Grundmann,Aachen

Friedrich Schönholtz †, Bad Hersfeld

Revised by Dipl.-Ing. (FH)Herbert Eidam, Bad Hersfeldand Dipl.-Ing.Bernd Rahn, Berlin

Elementary FanTechnology

The present „Fan Primer“ is aimedat contractors and operators.

Process equipment today wouldbe inconceivable without fans andpumps. Fans are indispensable forconveying gas mass flows, andthey perform essential functions indiverse process environments. Abasic understanding of fan techno-logy is therefore vital for contrac-tor and operator. It is the intentionof this „Fan Primer“ to impart therequisite fundamentals of fluid dy-namics and technology as well asof key fan functions, designs andperformance characteristics in apractical application context. Theboundary conditions and perfor-mance limits of the individual fantypes are also examined.

To the fan manufacturer or desi-gner this publication will be of limi-ted use. It cannot, and is not inten-ded to, resolve any of the issuesaddressed in this highly speciali-zed industry. Users from thesefields are therefore referred to therelevant academic and trade litera-ture.

Over and beyond the issues tou-ched upon in this Fan Primer, TLTTurbo-GmbH’s engineers will beglad to provide assistance withany problems this book cannotsolve.

Table of contents

I. Introduction

1.1 What is a fan? . . . . . . . . . . . . . . . . . . . 2.21.2 Designs . . . . . . . . . . . . . . . . . . . . . . . . 2.3

II. Basic fluid dynamics

2.1 Fluid flow. . . . . . . . . . . . . . . . . . . . . . . 2.42.2 Altitude formula . . . . . . . . . . . . . . . . . . 2.42.3 State variables for ideal fluid flow/

Bernoulli’s law . . . . . . . . . . . . . . . . . . . 2.42.4 Continuity equation . . . . . . . . . . . . . . . 2.52.5 Pressure loss . . . . . . . . . . . . . . . . . . . 2.52.5.1 Pressure loss due to surface

friction drag . . . . . . . . . . . . . . . . . . . . . 2.52.5.2 Pressure loss due to form drag. . . . . . 2.72.5.2.1

Impact loss . . . . . . . . . . . . . . . . . . . . . 2.82.5.2.2

Diffusion loss. . . . . . . . . . . . . . . . . . . . 2.82.6 Characteristic curve of a system. . . . . 2.82.7 Bernoulli’s law for real fluid flow . . . . . 2.92.8 Velocity distribution in the pipe or duct . 2.92.9 Pressure measurements . . . . . . . . . . 2.10

III. Axial-flow fans

3.1 Structure and operation. . . . . . . . . . . 2.113.2 Velocity triangels . . . . . . . . . . . . . . . . 2.113.3 Axial-flow fan designs . . . . . . . . . . . . 2.133.3.1 Axial-flow fans for air-handling

applications . . . . . . . . . . . . . . . . . . . . 2.133.3.1.1

Guide vanes . . . . . . . . . . . . . . . . . . . 2.133.3.1.2

Impeller blade configuration . . . . . . . 2.133.3.2 Axial-flow fans for industrial uses/

axial blowers . . . . . . . . . . . . . . . . . . . 2.143.3.2.1

Axial-flow fan with adjustable impellerblades and fixed outlet guide vanes . 2.14

3.3.2.2Axial-flow fan with adjustable inlet guidevanes and fixed impeller blades . . . . 2.15

3.3.2.3Speed-controlled axial-flow fans . . . . 2.16

3.3.3 Airflow direction inside the fan . . . . . 2.173.3.4 Hub ratio . . . . . . . . . . . . . . . . . . . . . . 2.173.3.5 Drive type . . . . . . . . . . . . . . . . . . . . . 2.17

IV. Centrifugal fans

4.1 Structure and operation. . . . . . . . . . . 2.194.2 Velocity triangels . . . . . . . . . . . . . . . . 2.194.2.1 Backward curved blades. . . . . . . . . . 2.194.2.2 Backward inclined straight blades. . . 2.194.2.3 Radially ending blades . . . . . . . . . . . 2.194.2.4 Forward curved blades . . . . . . . . . . . 2.194.3 Centrifugal fan configuration . . . . . . . 2.204.3.1 Type designations . . . . . . . . . . . . . . . 2.204.3.2 Inlet types . . . . . . . . . . . . . . . . . . . . . 2.214.4 Types and drive arrangements . . . . . 2.224.4.1 Casing orientation and direction

of rotation . . . . . . . . . . . . . . . . . . . . . 2.22

4.5 Important custom andspecial designs . . . . . . . . . . . . . . . . . 2.23

4.5.1 Centrifugal plug-in fans . . . . . . . . . . . 2.234.5.2 Roof-mounting centrifugal fans . . . . . 2.244.6 Operation under dust

and wear loads . . . . . . . . . . . . . . . . . 2.264.6.1 Conveying dust and fibrous media . . 2.264.6.2 Fan wear . . . . . . . . . . . . . . . . . . . . . . 2.27

V. Fans as system components

5.1 Characteristic system/fan curves,proportionality law . . . . . . . . . . . . . . . 2.28

5.2 Dimensionless variables . . . . . . . . . . 2.315.3 Selection criteria . . . . . . . . . . . . . . . . 2.325.4 Parallel operation . . . . . . . . . . . . . . . 2.345.5 In-line/series operation . . . . . . . . . . . 2.345.6 Pressure measurement on fans . . . . 2.35

VI. Speed control

6.1 Throttle control . . . . . . . . . . . . . . . . . 2.386.2 Blade pitch control. . . . . . . . . . . . . . . 2.396.3 Blade pitch adjustment . . . . . . . . . . . 2.396.4 Inlet vane control. . . . . . . . . . . . . . . . 2.39

VII. Drive unit dimensioning

7.1 Motors . . . . . . . . . . . . . . . . . . . . . . . . 2.407.2 V-belt drive . . . . . . . . . . . . . . . . . . . . 2.407.3 Couplings . . . . . . . . . . . . . . . . . . . . . 2.40

VIII. Explosion protection on fans

8.1 Standards situations . . . . . . . . . . . . . 2.418.2 Product standard for fans . . . . . . . . . 2.428.3 Marking example. . . . . . . . . . . . . . . . 2.428.4 Design notes . . . . . . . . . . . . . . . . . . . 2.438.5 Explosion protection of fans,

illustrated for a direct-drivencentrifugal fan . . . . . . . . . . . . . . . . . . 2.43

IX. Installation and dimensioning notes

9.1 Free inlet . . . . . . . . . . . . . . . . . . . . . . 2.449.2 Free outlet . . . . . . . . . . . . . . . . . . . . . 2.449.3 In-duct fans . . . . . . . . . . . . . . . . . . . . 2.469.4 Parallel and in-series operation. . . . . 2.47

2

Elementary Fan Technology 2

I. Introduction

1.1 What is a fan?

A fan is a turbomachine convertingenergy into the fluid flow of a gaseousmedium. The purpose of a fan is toconvey a volume of a gaseous medi-um (usually air) through a system(unit). As the system resists the flow

of this medium, the fan must overco-me this resistance by generating apressure head (total pressure diffe-rence). It is usually the core machinein the system it serves.

The following key variables play a role in fan specifications:

Symbol Formula Dim. Name

V cm*A m3/s Volume flow

cm V/A m/s Mean velocity

A �/4 (Da2 - Di2) m3 Cross-sectional area

Da m Outside diameter

Di m Inside diameter

v Di/Da – Hub ratio

pt1 Pa Inlet pressure

�pt pt2 – pt1 o. �H · � Pa Total pressure difference

kg/m3 Density

� cp/cv – Exponent *.)

f – Compression factor *.)

H m Gas column head

Pfluid W Fluid power

P Pfluid/� W Shaft power

� Pfluid/P – Efficiency

n rpm Rotational speed

u � · D · n/60 m/s Blade tip speed

� cm/ua – Capacity coefficient

� – Pressure coefficient

1,2,a,i,m Indices

�ring surface area in the case of axial-flow fans!�

*) Neglected in ventilation and air-condition technics (�pt < 2500 Pa)

�

� – 1

p1�pt

p1+�pt

p1( )� � –1· ·

V · �pt · f

p

2 · �pt

Ua2 · �

·

·

� – 1�

�

· f*


2

1.2 Designs

The first and foremost objective ofevery fan manufacturer in dimensio-ning his product for a given applicati-on is to maximize its efficiency in or-der to reduce energy costs. Basically,there exist four fundamentally diffe-rent fan designs named according tothe direction of the flow line throughthe impeller.

a) Axial-flow fanA straight flow line extends axially through the impeller.

c) Semi-axial flow fan (Bifurcated fan)A hybrid between axial and centrifugal designs, this fan is characteri-zed by a curved flow line through the impeller.

b) Centrifugal fanA straight flow line extends radially through the impeller (vertical tothe fan axis.

d) Centrifugal fans without spiral casing (centrifugal plug-in fan)Its flow line extends in virtually the same direction as in a centrifugalunit with spiral casing.

2


II. Basic fluid dynamics

2.1 Fluid flow

The fluid conveyed by a fan is in itsgaseous state. In ventilation and air-conditioning systems, air is the con-veyed medium. Its characteristics aredescribed by several state variablesand material properties. The most im-portant state variables are given be-low.

Temperature Tmeasured in K (degrees Kelvin)Pressure p measured in Pa

The most important material pro-perties are the following:Gas constant Rmeasured in Nm/kg KViscosity v measured in m2/sDensity measured in kg/m3

The relationship between state varia-bles and material properties isexpressed by the gas equation:

The gas constant of air isR = 287 Nm/kg · K

The absolute temperature T starts at-273°C = 0 K

Accordingly,+20°C is equal to 293 K

From the above, the density of air at0°C and p = 101325 Pa (= 760 torr)can be calculated as

Pressure dependence of the air’sdensity is low enough to be neglec-ted, at least at the pressure differenti-als encountered in a ventilation andair-conditioning context. In otherwords, the air is deemed to be a „non-compressible“ medium.

Temperature dependence of the air’sdensity, on the other hand, needs tobe taken into account. According tothe gas equation, the following holdstrue for different temperatures at thesame density:

The stated reference valuesTO = 273 K (= 0°C) and 0 = 1,29kg/m3 give us an equation for calcu-lating the air density at x°C :

For example: What is the density ofair at 20°C?

Note:

The above values apply to dry air.The density of moist air is slightly lo-wer. However, this influence is gene-rally negligible.

2.2 Altitude formula

If a fan is to be installed not at sea le-vel but in the mountains at an altitudeH, the density of air at that altitudehas to be determined. By internatio-nal agreement, the pressure Pa at al-titude H is calculated as

where pao is the pressure at sea leveland H is the altitude (in meters) abo-ve sea level.

Density may then be determined forthe stated temperature according tothe gas equation.

2.3 State variables for ideal fluidflow / Bernoulli’s law

Flow of a fluid is described in terms ofvelocity, static pressure and elevati-on. These are the „state variables“which are interrelated according toBernoulli’s law.

Under this law, the sum of velocity,pressure and elevation energies areequal at any point of the flow (assu-ming stationary flow*)), i.e.

where:

= density in kg/m3

c = mean flow velocity in m/s

ps = static pressure in Pa

g = acceleration due to gravity= 9,81 m/s2

h = elevation in m

In the case of an airflow, the elevationterm of the ·◊g◊· h equation (i.e. theweight of the air column) can beneglected due to its marginal value.This gives us the following expressi-on:

is referred to as the velocity heador dynamic pressure pd, while thesum of the dynamic and static pres-sure is called total pressure pt.

*) A flow is deemed to be stationary if the statevariables do not vary with time at a given point.

�

= =======pR·T

101325287·273

= kg/m3 = 1,29 kg/m3

1 T0

T1= or =

T0

T1

= 1,29 kg/m3273273 + x

= 1,29 kg/m3 = 1,2 kg/m3273273 + 2020

pa = pao · 287 – 0,0065 · H287� �5,255

c2 + ps + · g · h = constant2

c2 + ps = constant2

pt = c2 + ps = pd + ps2�

0�

�

0�0��

1

�

x�

�

��

�

�

�

�

c2

�

2


2

Bernoulli’s law, in this form, statesthat total pressure is the same at anypoint of the flow. This may be illustra-ted by a simple example, viz. the flowof a medium through a duct of varyingcross-section.

2.4 Continuity equation

The second basic equation of interestin this context is the continuity equati-on. It states that in a system with asingle inlet and a single outlet (i.e. anunbranched duct), volumetric flow ra-te will be identical at all points.

where:

V̇ = volume flow in m3/s

c = flow velocity in m/s

A = cross-sectional area

2.5 Pressure loss

Unlike their ideal counterpart, realfluid flows are subject to pressure los-ses. In a real-life system, these los-ses must be added to the load whichthe fan is required to overcome. A di-stinction is made between two typesof resistance, or drag:

a) surface friction dragb) form drag (also referred to as

pressure drag)

2.5.1 Pressure loss due to surfacefriction drag

As its name implies, this is a pressu-re loss due to friction encountered bythe airflow. It is calculated as follows:

For circular tubes:

�p refers to a pressure difference - inthis case, it stands for the pressuredifference between two points of theduct set apart by a distance l.

For ducts of any cross-section:

where:

= friction coefficient (dimensionless)

l = duct length in m

d = duct diameter

dh = hydraulic diameter in m

A = cross-sectional area in m2

U = wetted circumference in m

Examples: a) Rectangular duct ha-ving the sides a and b.

V̇ = A1 · c1 = A2 · c2 und c2 = c1A1A2

V̇ = c · A = constant

ld

ldh

with dh = 4 AU

dh = =4ab2(a + b)

2aba + b

l(a+b)2ab�pv = pd

�pv = · · pd

�pv = · · pd

2


b) Circular duct having the diametersd1 and d2:

Values of are taken from diagrams,e.g. Moody diagrams. They dependon the roughness of duct walls and onthe Reynolds number

Re = of the flow.*

Special diagrams exist in which theabove relationships are already ana-lyzed and expressed for a 1-meter-long section of ducting. It is assumedthat the duct is circular. For rectangu-lar ducts, the same diagrams areused but the duct diameter d is repla-ced with the relevant hydraulic dia-meter:

* is the kinematic viscosity of the fluid. For air

at 20°C, = 15 · 10-6

d 1 d 2

m2

s

Pressure loss due to friction resistance (surface friction drag) in a straight andhydraulically smooth duct:

dh = = d2 – d14 (d2

2 – d12)

� (d1 + d2)

�4

�pv = pdl

d2 – d1

c · d

The above diagram of pressure los-ses per 1 m of ducting applies to hy-draulically smooth ducts. For ductswith a less smooth finish, the �pvo va-lue obtained from the diagram mustbe adjusted by determining the ductsurface roughness k from the table ofduct types, then obtaining the correc-tion factor Ck from the diagram below.

�pv = Ck · �pvo [Pa] per 1m of duct

Roughness k /m [mm ]

Duct type k

Plastic tubing 0,005

Asbestos cement tube 0,1

Steel pipe 0,1

Sheet metal duct 0,15

Flexible hose 0,7

Wooden ducting 2,5

Concrete ducting 0,8

Masonry ducts 4,0

For a duct with rough surfaces, it maythus be written:

Volume flow V.

[m3/h]

Pre

ssur

e lo

ss �

pvo

[Pa]

or

Ro

[Pa]

ove

r 1

m o

f duc

tC

orre

ctio

n fa

ctor

Ck

Pressure loss �Pvo [Pa/m]

example


2

2.5.2 Pressure loss due to formdrag

Pressure losses resulting from formdrag may be attributable to variouscauses, e.g. duct elbows or tees,changes in cross-section, valves, orcomponents such as air heaters, coo-lers, filters, etc.

Such pressure losses are calculatedby the equation

wherein � is referred to as the resi-stance (or drag) coefficient.

The appropriate values of � mustusually be determined experimentallyand will be provided by the compo-nent manufacturer.

An overview of key � values is givenbelow.

�pv = � · c2 = � · pd2

Source: Taschenbuch für Heizung und Klimatechnik [HVAC Technology Manual], Recknagel-Sprenger, 58th ed.

�

2


2.5.2.1 Impact loss

An important type of form drag whichcan be calculated with sufficient ac-curacy is the sudden deceleration ofthe flow which occurs where the ductexpands abruptly.

Pressure loss resulting from the dec-line in flow velocity from c1 to c2 is re-ferred to as impact loss. It may be de-termined via the following equation:

The � values for this impact loss areshown in Diagram 1 below. The resi-stance coefficient for a one-sidedduct expansion is given in Diagram 2.

2.5.2.2 Diffusion loss

When the change in cross-section oc-curs gradually instead of abruptly, adiffuser is said to exist in the duct.The function of a diffuser is to decele-rate the fluid flow, thus converting dy-namic into static pressure („pressurerecovery“). The efficiency of this con-version depends closely on the ope-ning angle �. When it exceeds 10deg., flow ceases to adhere to theduct wall. Flow separation or ‘stalling’is said to occur. This effect causesvery substantial losses.

The following diagram shows � valuesfor diffusers with various openingangles �.

2.6 Characteristic curve of asystem

The sum of all pressure losses occur-ring on a fan’s inlet and outlet side gi-ves the total pressure difference �pt

for a given volume flow V. Total pres-sure difference is an important fan di-mensioning and selection parameter.The value pair �pt and V also marksa point on the system’s characteristiccurve, which is sometimes referred toas its parabolic drag curve. Since withturbulent flow*) the losses are propor-tional to the square of the velocity orvolume flow, a parabolic square curveis obtained when �pt is plotted over V.When this parabolic curve is drawnon log-log paper, it becomes astraight line having the gradient 2. Bynow taking the logarithm of �pt = kV2,we get log �pt = 2 log V + log k whe-re k is a system-specific constant.

*In some elements, such as filters, flow may benon-turbulent (low-turbulence displacementflow). Such elements must be considered se-parately in the calculations.

Diagram 1

Diagram 2

Linear representation of a system’s characteri-stic curve

Logarithmic representation of a system’s cha-racteristic curve

�pv = � · (c2 –c2)2= � · c12 (1– )2

2

�

A2

A1

2

�

Design point x

Design point x


2

The linear graph has the advantageof appearing more familiar and there-fore easier to read. Intermediate va-lues can be quickly interpolated. Onthe other hand, changes in the sy-stem’s characteristics are easier toconstrue in the diagram on log-log pa-per, since all characteristic curvesform parallel straight lines having agradient of 2.

The parabolic curve for a given sy-stem need not necessarily passthrough the zero point of the �p -Vdiagram, but may also show the pat-tern illustrated in the following graph.This will be the case, e.g. if a fan isdelivering its output into an overpres-sure chamber or pressure vessel. Itspressure difference against the at-mosphere is �p1. The system’s cha-racteristic curve will then intersect thevertical �pt axis at the point �p1.

2.7 Bernoulli’s law for real fluidflow

By inserting the loss terms for surfacefriction and form drag, Bernoulli’s lawcan be extended to apply to real fluidflow. The following will then hold truefor two points (1) and (2) of a flow ifthe elevation term is neglected:

where

and

c12 + p1 = c2

2 + p2 + �i· pdi

+ · · pdi2 2 n

i = 1

n

i = 1

m

i = 1

lidi

lidi

2.8 Velocity distribution in the pipeor duct

Due to surface friction and flow adhe-sion to the duct walls, the velocity dis-tribution across the duct diameter isnot constant. Instead, a so-called ve-locity profile can be observed. Onlydownstream of an inlet nozzle flow isalmost homogeneously distributed.Once it has passed a certain down-stream length of ducting, the profilehas formed.

Formation of this velocity profile mustbe duly taken into account, particular-ly in measurements aimed to determi-ne, e.g. volumetric flow rates.

Distorted velocity profiles and irregu-lar pressure distributions across theduct diameter will occur downstreamof in-duct baffles, obstacles or deflec-tion points. Duct elbows or curves aregood examples of this phenomenon.

Downstream of the deflection pointthe medium becomes detached fromthe walls, which results in a highly ir-regular velocity profile along the insi-de of the duct. Moreover, static pres-sure is higher on the outside than to-ward the center, where negative pres-sures may actually occur. This effectcan be greatly diminished by instal-ling baffles, which will also reduce theresistance (or drag) coefficient (referto section 2.5.2).

�i · pdi = sum of all (n) form drag influences between the points (1)and (2),

m

i = 1 · · pdi = total of all (m) surface friction influences between the points

(1) and (2)

�p t

ød

10d

Velocity profile re-turns to a balancedstate after approx.6dh

dh = hydraulic diameter

� �

2


2.9 Pressure measurements

The following sketches illustrate fun-damental options for measuring pres-sures ps, pd and pt.

ps static pressure, i.e. pressure actingon a wall parallel to the direction offlow

pd dynamic pressure, or velocity head

pt total pressure, i.e. sum of static anddynamic pressures

Measurement on outlet side

Measurement on inlet side

� Static pressure ps is measured bymeans of a pressure gauge via acarefully deburred orifice in theduct wall. Best results are obtainedby providing several such orificesalong the circumference inter-connected via a ring line.

� Total pressure pt can be measuredwith a 90° angle probe held fron-tally into the oncoming flow. Suchprobes are referred to as Pitot tu-bes.

� Dynamic pressure is determinedas difference between pt and ps.From pt = ps + pd, it follows that pd

= pt - ps

A device commonly used for dyna-mic pressure measurements is thePrandtl tube, which combines a Pi-tot tube with the functions of a sta-tic pressure probe.

To perform measurements within asystem, it is best to select a pointwhere a uniform velocity profile pre-vails. Measuring locations immedia-tely downstream of elbows (refer tosection 2.8), t-fittings or diameter ex-pansions should be avoided sincestatic pressure will not be constantacross the duct diameter here andmeasurements will necessarily be fla-wed.

Today, standard pressure gauges willnormally show pressures in Pa. Olderdevices may still give readings in mm-WC (millimeters water column). 1mmWC = 1 kp/m2.

Conversion into the applicable sy-stem (SI units) is made according tothe following formula:

1 mm WS = 1 kp/m2 = 9,81 Pa � 10 Pa

� � �

� � �

ambient pressure


2

III. Axial-flow fans

3.1 Structure and operation

An axial-flow fan consists of bell-mouth built into the casing, impeller,drive motor, and assembly of outletguide vanes (or, in the case of axial-flow fans without outlet guide vanes,motor mounting bracket).

Large axial-flow fans are equippedwith a diffuser on the outlet side toachieve a low-loss conversion of thehigh dynamic head into static pressu-re. Diffuser designs may vary, depen-ding on whether or not the fan has anoutlet guide system.

The purpose of the bellmouth is toproduce a uniform velocity distributi-on in front of the impeller so that theimpeller vanes will be exposed to theflow over their full surface area (referto section 2.8). The conversion ofenergy takes place in the impeller bla-de channels. Both static and dynamicpressure is produced here. Down-stream of the impeller the flow is in-tensely turbulent and swirling, i.e. theairflow exiting the impeller has a tan-gential velocity component.

To convert this useless component ofdynamic pressure energy into its sta-tic equivalent, guide vane systemsare employed. These vanes are ar-ranged as a stationary ring in theshaft, either downstream or upstreamof the impeller. Depending on theirposition, they are referred to as inletor outlet guide vanes. They deflectthe flow so that it will exit in an axialdirection from the fan.

Casing

Diffuser(recommended option)

Impeller

Bellmouth Outlet guide vanes

Motor bracket

Impeller without outlet guide vanes

Motor bracket

Motor

3.2 Velocity triangles

Flow conditions inside the fan can begraphically represented by means ofvelocity triangles. In these triangles,the following symbols and indexesare used:

Index 0 Entry into inlet guide vanesIndex 1R Entry into impeller or exit

from inlet guide vanesIndex 2 Exit from impeller or entry

into outlet guide vanesIndex 3 Exit from outlet guide vanes

c Absolute velocityw Relative velocityu Impeller blade tip speed (circum-

ferential velocity)

The absolute flow velocity c always isthe vectorial sum of tip speed u andrelative flow velocity w:

c1R is the swirl-free absolute entry ve-locity into the impeller (� note thering cross-section).

c = u + w� � � Impeller direction of rotation

ImpellerW1 c 1R

Bladeprofile

2


u is the peripheral impeller velocity(blade tip speed), which is related tothe fan’s rotational speed (rpm) ac-cording to the following function:

where

� = angular velocity tip speed of theimpeller in s–1

u = peripheral velocity in m/s

d = diameter of blade cross-section in m

n = impeller rotational speed in rpm

w1 = relative velocity of approach flowon the blade. This variable is ob-tained by vectorial addition of inletvelocity c1 and peripheral velocityu, wherein the length of the vec-tors is equivalent to the amount ofthe velocity.

Change from w1 to w2 is a result of thecurvature and shape of the bladechannels.

c2 is the absolute velocity at the exit ofthe blade cascade and hence, at thepoint of entry into the outlet guide va-nes.

a) Axial-flow fan without guide vanes

b) Axial-flow fan with outlet guide vanes

c) Axial-flow fan with inlet guide vanesInlet guide vanes

Bell-mouth

Impeller

d) Counter-rotating axial flow fans

To boost pressure output, axial-flow fans can sometimes be used in pairs ofcounter-rotating units. Such a configuration requires two complete fans,each having its own motor, which are installed with their (counter-rotating)impellers immediately facing each other.

A counter-rotating fan system does not differ significantly in aerodynamicterms from a two-stage co-rotating fan configuration, although acousticemission levels are much higher in the case of the former.

u 2=

u

u 1=

u

Impe

ller d

irect

ion

ofro

tatio

n

w 2

c2

w 1c1R

u 1=

u

w 1

c1R

co

u 1=

u

w 1

c 1R

c3 = c1RImpe

ller

dire

ctio

nof

rot

atio

n

Impe

ller

dire

c-tio

n of

rot

atio

n

Inlet guide va-nes (stationary)

Inlet guide vanes(stationary)

u 2=

uw 2

c2 c 2

u

u 2=

u

w2

c2Section AB

A B

ød

Motor

Motor bracket

Impeller

Bell-mouth

Motor

Motor

Casing

Casing

Impeller

Bell-mouth

Motorbracket

Outletguidevanes

u = · � = d2

d · � · n60

Motor bracket

Casing


2

3.3 Axial-flow fan designs

Axial-flow fans can be classified ac-cording to diverse application andoperating criteria.

3.3.1 Axial-flow fans for air-hand-ling applications

3.3.1.1 Guide vanes

� Axial-flow fan without guide vanes� Axial-flow fan with inlet guide va-

nes� Axial-flow fan with outlet guide va-

nes

3.3.1.2 Impeller bladeconfiguration

Axial-flow fans with fixed, non-adju-stable impeller blades have only oneconstant characteristic curve for eachrotational speed.

Axial-flow fans with pitch-adjustableimpeller blades have multiple charac-teristic curves plotted as a function ofthe blade angle. They offer the ad-vantage of being particularly adapta-ble to diverse operating conditions.

In a standard design with outlet guidevanes impeller blades are pitch-adju-stable when the fan is stationary. Forstraightforward air-handling applicati-ons (i.e. low pressures), units withoutoutlet guide vanes but with stationaryimpeller blade adjustment are alsoused.

Example:Axial flow fan (blade pitch adjustableon stationary fan)

Manufacturer & type:TLT-Turbo GmbHType AXN 12/56/800/M-D

Type M-D

Am Weinberg 68 · D-36251 Bad Hersfeld/GermanyTel.: +49.6621.950-0 · Fax: +49.6621.950-100

Volume flow [m3/h] or [m3/s]Dyn. pressure [Pa] or x0.1 [kp/m2]Flow velocity [m/s]

Blade tip velocity u2 = 60 m/sTemperature t = 20°CDensity = 1,2kg/m3

Moment of inertia l = 0.69 kg m2

Int.casing diameter 797 mmOutlet cross-section A2 = 0.5 m2

CHARACTERISTIC CURVES OF AXIAL-FLOW FANSWITH DIRECT DRIVE AND OUTLET GUIDE VANESTYPE AXN 12/56/800D*ROTATIONAL SPEED 1450 RPM

Shaft power inputrequirement

with 2.5 D duct

free outlet

Max. availablemotor sizes:

refer to dimen-sional sheets

Airflow direction D (outlet over motor) - airflow direction S (inlet over motor) available upon request - values rounded to standard figures.

Tot

al a

cous

tic p

ower

leve

l

Blade angle

Tot

al p

ress

ure

incr

ease

�p t

[Pa]

→

Pw = =[kW] V · �pt

� · 1000 · 3600

2


3.3.2 Axial-flow fans for industrialuses / axial blowers

For practical purposes, this fan cate-gory is subdivided into the followingtypes:

3.3.2.1 Axial-flow fan with adjusta-ble impeller blades and fixed outletguide vanes

Such axial-flow fans are available

� with individually adjustable impel-ler blades, adjusted on the statio-nary fan

� with centrally adjustable impellerblades, adjusted on the stationaryfan

� with jointly controlled impeller bla-des, adjusted under load (i.e. whilethe fan is running). This design of-fers certain advantages in control-ling volume flows and provides avery broad operating range withgood part-load characteristics.

Hydraulic blade pitch adjustment un-der load is now state-of-the-art tech-nology.

Example:Axial-flow fan with impeller bladepitch adjustment

Manufacturer:TLT-Turbo GmbH

Fan casing - top part

Dual-stage rotor

Coupling halves

Compensator

Diffuser

Fan casing - bottom part

Inlet chamber

Hydraulic adjustment mechanism

Deflector

Blade pitch adjustmentactuator

Oil supply system

Bearing temperature indicator

Axial-flow fan with hydraulic blade pitch adjustment under load

Dis

char

ge h

ead

�m g

as c

olum

n�

Volume flow V �m3/s�

10000

9000

8000

7000

6000

5000

4000

3000

2000

1000

00 100 200 300 400 500 600 700 800 900 1000 1100 1200

� = %

�88

86

83

80

75

70

60

50

40

Acousticinsulation

Intermediate shaft

Anti-vibration mounts


2

3.3.2.2 Axial-flow fan with adjusta-ble inlet guide vanes and fixed im-peller blades

The part-load performance of this fantype is usually inferior to that of axial-flow units with adjustable impeller bla-des.

However, given their rugged design,these fans are preferred for use undersevere operating conditions, e.g. inhigh-temperature or high-dust envi-ronments.

Typical applicationsPower stations, mining

ExampleAxial-flow fan with adjustable inletguide vanes

Manufacturer:TLT-Turbo GmbH

Axial-flow fan with inlet guide vanes

Dis

char

ge h

ead

�m g

as c

olum

n�

Volume flow V �m3/s�

10000

9000

8000

7000

6000

5000

4000

3000

2000

1000

00 100 200 300 400 500 600 700 800 900 1000 1100 1200

� 87,587

8582

7974

63

53

42

31

20

10

2


3.3.2.3 Speed-controlled axial-flowfans

Frequency converters have evolvedinto a powerful means of controllingthe rotational speed of electric mo-tors. This makes them ideal for usewith fans.

Especially axial-flow fans with indivi-dual impeller blade adjustment on thestationary unit benefit from the use ofadvanced frequency converter tech-nology for motor rpm control. Advan-tages are manifold:

� favourable placement of the axial-flow fan’s operating point on thecharacteristic curve

� very good part-load performancegiving a square-law characteristiccurve for the system

� favourable acoustic properties inpart-load operation

� simple mechanical structure ensu-res trouble-free operation

Example:Axial-flow fanSpeed controlled (impeller bladesadjustable on stationary fan)

ManufacturerTLT-Turbo GmbHType AXN 12/56/1400/R2

Type M-D

CHARACTERISTIC CURVES OFAXIAL-FLOW FANS WITH BELT DRIVETYPE AXN 12/56/1400/RSPEED CONTROLLEDAm Weinberg 68 · D-36251 Bad Hersfeld/Germany

Tel.: +49.6621.950-0 · Fax: +49.6621.950-100

Characteristic curves shown below apply to a 23°blade angle.Temperature t = 20°C, density � = 1.2 kg/m3

Number of blades: 12Moment of inertia l = 10,05 kg/m2

Int. shaft diameter: 1415 mmOutlet cross-section A2 = 1,57 m2

These characteristic curves were measured with 2,5D ducting on fan outlet. Efficiencies apply to max.rpm

Type R1 notavailable

Type R2max. 90 kW

App

rox.

sha

ft po

wer

inpu

tre

quire

men

t Pw

[kW

]

Fan

rpm

Bla

de ti

p ve

loci

ty u

[m/s

]

Tot

al a

cous

tic p

ower

le

vel L

w[d

B]

Tot

al p

ress

ure

incr

ease

� p

t[P

a]→

Max

. ava

ilabl

e m

otor

siz

es:

refe

r to

dim

ensi

onal

she

ets

Volume flow V.[m3/h]

Volume flow V.[m3/h]

Flow velocity c1 = c2 [m/s]

Dynamic pressure pd [Pa]

values rounded to standard figures.


2

3.3.3 Airflow direction inside the fan

Airflow in a fan commonly passesfrom the impeller and guide vanesover the motor and bearing assemb-ly. All characteristic curves are basedon this layout.

However, process reasons may re-quire an arrangement of the motor onthe fan inlet side. For these applicati-ons TLT-Turbo GmbH provides „inletover motor“ (S) type units.

Nevertheless, the „D“ airflow directionshould be preferred since „S“ typefans require a devaluation of the cha-racteristic curve and achieve inferiorefficiency levels.

3.3.4 Hub ratio

The hub ratio denotes the ratio of theimpeller hub diameter to the externalimpeller diameter. In the case of axi-al-flow fans, this ratio commonly va-

ries between 0,25 and 0,63. By com-parison, axial-flow compressors mayhave larger hub ratios.

The smaller the hub ratio, the lowerthe pressure of an axial-flow fan.

3.3.5 Drive type

Standard designModel AXN, type M-D(outlet over motor)

Special designModel AXN, type M-S(inlet over motor)

Axial-flow fan - standard direct-drive type Type M - Impeller on motor output shaft

Axial-flow fan - V-belt driven type(motor mounted sepertely on base-frame)

Type R2 - Impeller driven via V-belt

Axial-flow fan - V-belt driven type (motor mo-unted on fan casing) for light air-handling duty

Type R1 - Impeller driven via V-belt

2


Large axial-flow fan (blower) - dualstage design with a common doublebearing, driven directly via a couplingand intermediate shaft. The electricmotor is arranged outside the gasflow.

Horizontal installation!

Large axial-flow fan (blower) - singlestage with double bearing, driven di-rectly via a coupling and intermediateshaft. The electric motor is mountedvertically outside the gas flow.

Vertical installation!e.g. in a stack

Large axial-flow fan (blower) - singlestage, impeller mounted on the motorshaft, electric motor arranged in thegas flow.

Vertical installation!

DiffuserElectricmotor

Inlet nozzle

Maintenance space

Maintenance space


2

IV. Centrifugal fans

4.1 Structure and operation

A centrifugal fan has a spiral casingwith bellmouth and an outlet connec-tion, impeller, and discharge cut-off.The airflow enters the impellerthrough the bellmouth and is deflec-ted centrifugally. A conversion ofenergy takes place within the impeller(blade channel), i.e. the mechanicalenergy imparted to the impeller viathe shaft from the motor is transfor-med into pressure and velocity ener-gy. Functions of the spiral casing aretwofold. On the one hand, it gathersthe air exiting the impeller and guidesit to a common outlet. On the other, itconverts part of the velocity energy(dynamic pressure) into pressureenergy (static pressure) through thesteady expansion of its cross-section

in the direction of flow (diffuser ef-fect).

The narrowest point between casingwall and impeller is formed by the cut-off.

Centrifugal fans can deliver higherpressures than their axial-flow coun-terparts since their radial blade chan-nels promote the build-up of staticpressure through the different peri-pheral speeds at the impeller inletand outlet.

Bellmouth

Motor

Cut-off

Spiral casing

4.2 Velocity triangles

Centrifugal fans are classified intofour different impeller types accordingto the shape of their blades.

4.2.1 Backward curved blades

Centrifugal fans with backward cur-ved blades are also referred to as„high-performance“ fans due to theiroutstanding efficiency. These impel-lers are particularly suitable for plug-in fans.

Blade outlet angle w2 � 30°

4.2.2 Backward inclined straightblades

Such impellers are suitable for gasescontaining coarse dry particulate matter.Their efficiency is still very high, warran-ting classification in the high-performan-ce category. Centrifugal fans with thisblade configuration may be used tohandle dirty media or to convey materials(„high-performance dust fans“).

Blade outlet angle w2 = 40 to 60°

4.2.3 Radially ending blades

Such impellers are rarely employed ina ventilation and air conditioning con-text. Since the blade geometry reliab-ly prevents accretions, centrifugalfans of this type are used to conveygases containing high loads of dustand suspended particulates (pneu-matic conveyance applications). Ho-wever, depending on dust type, back-ward curved blades may also servethis purpose.

Blade outlet angle w2 = 75 to 90°

4.2.4 Forward curved blades

Centrifugal fans with many forwardcurved blades are also referred to asdrum rotor fans. The proportion of ve-locity energy obtained with this de-sign is very high. Due to the low effi-ciency achieved, use of such impel-lers is now limited to small centrifugalfans for air-handling applications.

u2

c2c1

w1

u1

w2

u2

u1

c2

c1

w2

w1

c2 w2

w1

u1

c1

u2

c1

u1

u2

w2c2

w1

Impeller

2


4.3 Centrifugal fan configuration

Centrifugal fans are habitually classi-fied according to the following criteria:

� Blade shape

a) Centrifugal fans with backwardcurved blades („high-performancefans“)

b) Centrifugal fans with backward inc-lined straight blades („dust fans“)

c) Centrifugal fans with radially en-ding blades for dirty industrial gasflows

d) Centrifugal fans with forward cur-ved blades for ventilation and air-conditioning (refer also to section4.2).

� Impeller characteristics

One important parameter is the ratiobetween the outside diameter and theinlet diameter (= nominal diameter) ofthe centrifugal impeller. This ratiocharacterizes the centrifugal fans in agiven range. Typical diameter ratiosvary between 1,1 and 7,1. In ventilati-on and air-handling applications, se-ries 11 and 14 fans are common. Thelarger the diameter ratio, the higherthe pressure delivered by the fan.

The centrifugal fan range of TLT-Tur-bo (formerly Babcock BSH) is structu-red into seven series delivering thefollowing pressures:

4.3.1 Type designations

Type designation of a centrifugal fanshould indicate not only its pressureoutput capability but also its specific

application properties. Apart from thefan series (reflecting the diameter ra-tio), this identification need is fulfilledby the blade outlet angle w2. As a re-sult, each fan series comprises va-rious impeller blade configurationsdefined by the blade outlet angle w2.The fan can thus be adapted individu-ally to specific application require-ments.

� Steep or flat characteristic curve

� Control range requirements

� High-dust service

� Wear or accretions

� Direct motor drive for individualoperating point selection

Type designation of TLT-Turbo GmbH’s standardrange of industrialcentrifugal fans

TLT-Turbo GmbH’s standard range isdivided into seven centrifugal fan se-ries, each comprising various bladeshapes and blade outlet angles.

In addition, each type can be made ofdifferent materials to resist chemicalattack and elevated temperatures.

Series Pressure range at = 1,20 kg/m3

(guide values)

11 100 – 2800 Pa14 1800 – 4500 Pa18 2800 – 7100 Pa22 5500 – 11200 Pa28 8100 – 16000 Pa35 12500 – 20000 Pa45 16000 – 25000 Pa

�

Diameter ratio 1,4 = Series 14

14 / 45

Series(Diameterratio x10)

Blade outletangle w2


2

The illustration across shows all typesin our standard range, together withtheir key properties. This product di-versity allows us to address each ap-plication requirement in an ideal man-ner.

� = Steep characteristic curve, ma-ximum efficiencies for industrialenvironments, particularly fa-vourable control response

� = For dust service, dust repellent,for coarse and dry suspendedparticulates

� = For extremly high dust loads,featuring self-cleaning impellerblades except for deposits dueto chemical reactions or elec-trostatic charge

Fan typespreferred inventilationand airhandlingapplications

Single-inlet centrifugal fan impeller

Double-inlet centrifugal fan impeller

11/20 �11/25 �11/30 �

11.1/30 �11/40 �

11/45 �11/60 �

14/20 �14/30 �14/45 �

14/60 �14/80 �

18/30 �18/50 �18/80 �

22/40 �22/55 �22/80 �

28/40 �28/60 �28/75 �

35/45 �35/75 �

45/50 �45/78 �

4.3.2 Inlet type

Centrifugal fans may be of the single-inlet or double-inlet type. A double-in-let centrifugal fan delivers approxima-tely twice the volume per unit timewhen compared to a single-inlet unitof the same nominal size and totalpressure increase. The configurationcorresponds to a parallel arrange-ment of two fans (refer to section 5.4).

2


4.4 Types and drive arrangements

4.4.1 Casing orientation and direc-tion of rotation

Housing orientation and direction ofrotation are always specified as vie-wed from the drive side.

For designations used, refer to theabove table.

Type examples(shown with options)

Type RUM: single-inlet, impeller on motor shaftend

Type RUR: single-inlet, belt-driven impeller

Type ZER: double-inlet, belt-driven impeller

Type RUK IV: single-inlet, direct driven via anelastic coupling

Type RUK V: single-inlet, direct driven via anelastic coupling

Type ZSKI: double-inlet, with inlet box, directmotor driven

Type Connection Drive

R U M

Z E K

S R

Single-inlet

Double-inlet

Direct ductconnection

Withbellmouth

With inletbox

Impeller onmotor shaft

viacoupling

via belt

* Design types according to VDMA 24164


2

4.5 Important custom and specialdesigns

4.5.1 Centrifugal plug-in fans

Configured preferably as a single-in-let unit, this fan type is preferred whe-re large volumes of air must be con-

veyed against total pressures≤ 2000 Pa.

Typical applications therefore inclu-de

Dryers (all types)Spray-painting linesCooling installationsCleanroom systemsCentral air-handling units

Centrifugal plug-in fan for installa-tion in a dryer

Driven by a standard motorMax. temperature: 250°C

Centrifugal plug-in fan for horizon-tal installation in central AHUplants

Driven by a standard motor mountedin the airflow

Centrifugal plug-in fan for verticalinstallation

Driven by a standard motor mountedin the airflow

2


4.5.2 Roof-mounting centrifugalfans

Centrifugal fans for rooftop installati-on are special free-inlet units suitablefor use as central air exhaust fansdue to their pressure capacity.

These fans are available in diverse ty-pes:

centrifugal roof fanDRH type

with horizontal air outlet,driven by a special motor(external rotor)

centrifugal roof fanDRV type

with vertical air outlet,driven by a special motor(external rotor)

centrifugal roof fanDRVF type

with vertical air outlet, driven by astandard motor


2

centrifugal roof fanBVD type

vertical air outlet, designed as asmoke exhaust fan to extract fumesand smoke, rated for 400°C/620°C -120 minutes

centrifugal roof fanDR-SDH type

with horizontal air outlet, noise-insu-lated on inlet and outlet side

centrifugal roof fanDR-SDV type

with noise-insulated vertical outlet

2


4.6 Operation under dust and wearloads

For exhaust air fans and some indu-strial process fans, dust and wear arefactors which require special conside-ration at the design and dimensioningstage. The dust load encounteredand its consistency and moisture areimportant criteria.

4.6.1 Conveying dust and fibrousmedia

Backward curved blade

Dust sticks tosurface.R > T

Conditionallysuitable for drydust

�FN FZ

RT

�

Radially ending bla-des

Dust is flungaway from bladesurface.R < T

For dirty indu-strial media

�

Impeller without coverplate

Fibrous media -glideover blade sur-faceR < T

Specifically for pneumaticconveyance of fibrous mat-ter!

F N

FN

F Z

FZ

RT

R

T

�

(Stationary cover plateattached to housing)

Explanation of terms

FN = Force in normal direction

FZ = Centrifugal force

T = Force in tangential direction

R = Friction force = FN ·µ

µ = Friction coefficient

For further information on how toselect suitable centrifugal fans re-fer to chapters 4.2 and 4.3

Every dust particle that does notadhere to a surface is a potentialcause of wear. While a lack of in-formation about the wear processwill primarily affect the question ofspare part availability for the selec-ted fan types, uncertainties concer-ning dust adhesion characteristicswill often determine whether or nota given fan is employed at all.

The tendency of suspended solidsto adhere on the blade inlet sides ofcentrifugal fan impellers with back-ward curved blades and on the bla-de outlet surfaces of forward cur-ved blades can only be avoidedwith any degree of certainty if theapplicable angles of slip are accu-rately known for the given dust par-ticle size distribution [1].

Note:High dust loads in the conveyedmedium require an additional po-wer input which must be taken in-to account!

Important:

With gas flows containing highdust loads, the resulting extrapower requirement and pressureloss must be taken into account.


2

4.6.2 Fan wear

Fans conveying media which containsuspended particles are subject towear. This effect can be reduced, al-beit not avoided altogether, throughsuitable design strategies.

Abrasive wear changes the surfacesexposed to the gas flow. Symptomsinclude denting, corrugation effects,scratches and score marks on the ex-posed metal. A micro-level „machi-ning“ process is taking place, resul-ting in a loss of material.

Abrasion is caused by particulatematter in the gas flow which slidesalong the relevant surfaces or collideswith them from various angles.

Abrasive processes and their termi-nology are addressed in DIN 50320.

The most important wear parameterscan be summarized thus:

A. Impeller– Hardness and material thickness of

the impeller body– Blade tip velocity– Blade shape

B. Dust load– Hardness of the impinging particles– Grain size and geometric particle

shape– Particle density

Wear processes

The influence of particle hardness onthe rate of abrasion from a soft surfa-ce (e.g. non-armoured blade) or ahard surface (e.g. hardfaced blade) isillustrated by the following diagram:

1 If the attacking particles are softerthan the exposed component, littleabrasion occurs. The process re-mains in the low wear range.

2 If the attacking particles are harderthan the exposed component, signi-ficant abrasion will take place. Theprocess lies in the high wear ran-ge.

3 If the hardness of the attacking par-ticles and of the exposed compo-nent are approximately equal, mi-nor shifts will suffice to produce asubstantial change in wear beha-viour. The process lies in the rangeof the steep rise.

Important

To minimize wear, the hardness ofthe exposed component must beselected such that it exceeds that ofthe abrasive particles.

Note:Anti-wear measures on impellers will give rise to increased weights and imbalanceforces. Consequences such as– need for reinforced driveshafts and bearings– need for stronger fan supporting structures– efficiency deteriorationneed to be taken into account!

The general principle whereby acentrifugal fan blade extending at atangent to the dust flow at everypoint of the blade's radial extensionwill always be subject to the leastamount of wear (i.e., sliding wear)can be considered proven. Wherea problem cannot be addressed byselecting appropriately adaptedblading, the engineer is left with theoption of maximizing economic effi-ciency via the selection of suitablematerials and material thicknesses.

Hardness of attacking particles

High wear

Low wear

Stee

p ris

e

Abr

asio

n ra

te �

�

� Soft component� Hard component

Measures Descriptions b

a1a2

a3a4

a5

a6

b = lateral protection

1. Blade material s Ste 70

2. Blade thickness „s“ increased by2-3 mm

3. Weld beads extending in a direc-tion transverse to the direction offlow, placed with the aid of hardfa-cing electrodes. Bead distance „a“decrease toward the outside dia-meter.

1. Blade base material s

2. Surface hardfaced to s1 = approx.0,8 – 1,0 mm by tungsten carbideflame spraying

1. Blade base material s

2. Surface hardfaced to s1 = approx.0,5 mm by continuous weld clad-ding with a material containingchromium carbide

s s 1

Flat blade(no curvature)

b

s s 1

Flat blade(no curvature)

b

2


V. Fans as systemcomponents

5.1 Characteristic system/fan cur-ves, proportionality law

Theory of establishing a system’scharacteristic curve was examinedearlier in section 2.5. Below we shalltake a look at the underlying laws byexamining linear and log-log graphsfor the example of a RA 11.1 centrifu-gal fan, nominal size 800, made byTLT-Turbo GmbH.

If two operating points are compared,pressure ratio is equal to volume ratiosquared, i.e.

In our example, the operating pointB1 lies at V̇1 = 10 m3/s and�pt1 = 1750 Pa. Which value is obtai-ned with�pt2 at ·V2 = 5 m3/s

�pt2 = 1750 Pa · = 438 Pa.( )2

510

�pt1�pt2

= or �pt2 = �pt1 ·( )2V1

V2( )2V2

V1

System characteristic curves withdifferent operating points

The total pressure increase producedby a fan consists of a static and dyna-mic component. The dynamic pressu-re increase is expressed with referen-ce to the fan inlet connection. It is cal-culated according to the known for-mula

where c is the mean flow velocity inthe fan inlet connection, i.e.

In our example, we obtain the follo-wing for

.V = 10 m3/s and the selected

NG 800 centrifugal fan:

Dynamic pressure in the fan inletconnection

(C = line of dynamic pressure)

pd = c2

pd = · c2 = · 19,92 = 238 Pa

c = ,where A is the cross-sectionalarea of the inlet connection.

V̇A

A = = = 0,502 m2d2 �4

0,82 m2 �4

c = = = 19,9 m/s10 m3

0,502 m2 · sV̇A

1,22

kg m2

m3 s2

2

2

·

·

·

·

Linear log-log

A = System characteristic curve B = Operating point

�

�


2

The performance behaviour of a fanis described by its characteristic cur-ve. This graph is determined by rig-testing under specific conditions defi-ned in DIN 24163. To establish thecurve, various operating points are si-mulated by throttling the volume flow,and the measured value pairs for �pt

- V are plotted in a diagram fromwhich the characteristic curve is thendrawn. During rig testing, shaft powerinput requirement is measured at thesame time to determine the fan’s effi-ciency. The power input requirementis obtained from the input torque MW

and the angular velocity ω. The effi-ciency h is the quotient of input andoutput power. The output P is referredto as the useful or effective power; thepower input is the shaft power requi-rement Pw.

P = �pt · V̇

Pw = MW · �

� = =P

PW

�pt · V̇

MW · �

�pt · V̇�

�Hence, Pw = =

if � is known.

P�

P = power in W (or kW if p1 is ex-pressed in kPa)

�pt = total pressure increase in Pa(or kPa, respectively)

V̇ = volume flow in m3/h

Mw = input torque in Nm

� = angular velocity in 1/s

� = · s–1 for n in rpm� · n30

Characteristic curve of fan and system

The fan’s operating point within theoverall system always lies at the in-tersection of the characteristic curvesof the system and the fan.

The point of intersection between thefan’s characteristic curve and the dy-namic pressure line marks the maxi-mum capacity, i.e., the air volumewhich this fan would deliver against„zero“ system resistance.

�

2


Proportionality laws

1) Rotational speed change (from n1

to n2, in our case from 1400 to1800 rpm)

In our example, the fan speed waschanged from 1400 to 1600 rpm

Given the known square law of thecharacteristic curve, this results in thefollowing changes:

a) Volume flow V changes in proporti-on to the speed (rpm), i.e.

b) Total pressure increase �pt chan-ges with the square of the rotatio-nal speed, i.e.

c) Shaft power input requirement PW

changes with the third power of therotational speed, i.e.

= or ·V2 = ·V1 ·

= or �pt2 = �pt1 ·

·V1·V2

�pt1

�pt2

n1

n2

n1

n2

n2

n1

n2

n1( )2 ( )2

= or Pw2 = Pw1 ·Pw1Pw2

n1

n2

n2

n1( )3 ( )3

Change in rpm (from n1 to n2, i.e., from 1400 to 1600 rpm in this example):

Proportionaltiy laws for fan seriesof geometrical and kinematicallsimilarityIndex 2 = Reference Size

Formular Symbols:·V = Volume flow [m3/h or m3/s resp.]n = Rotational Speed [rpm]�pt = Total pressure differences [Pa]Pw = Power requirement at shaft [kW]T = Temperature [°C]

= Density [kg/m3]d = Outer dia. of impeller Ø [m]

�

A n � const., = const.

B n = const., � const. bzw.

T � const.

V1 = V2 = const.

�

�C n = const., d2 � const.

D n � const., d � const., � const.�

linear log-log

=·V1·V2

n1

n2

�pt1

�pt2( )2n1

n2= =

·V1·V2

( )2

Pw1Pw2

( )3n1

n2= =

·V1·V2

( )3

�pt1

�pt2= =1

2

�

�T1

T2

Pw1Pw2

= =1

2

�

�T1

T2

=·V1·V2

d1

d2( )3

�pt1

�pt2=

d1

d2( )2

Pw1Pw2

=d1

d2( )5

=·V1·V2

n1

n2

d1

d2( )3

�pt1

�pt2=

n1

n2( )2 1

2

�

�d1

d2( )2

=n1

n2( )3 1

2

�

�d1

d2( )5Pw1Pw1


2

5.2 Dimensionless variables

To facilitate the assessment andcomparison of fans with regard totheir suitability for individual applicati-ons, dimensionless variables havebeen defined for key properties:

a) Efficiency

(Refer to Fig. 5.1)

with �pt in Pa, V in m3/s, and Pw in W.Efficiency � denotes the ratio of thefan’s power output to the requiredshaft power input. It thus measuresthe quality of the energy conversionprocess performed by the fan.

b) Pressure coefficient

with �pt in Pa, � in kg/m3 and u2 inm/s. Coefficient ψ measures the totalpressure difference delivered by a fanat a given blade tip velocity.

c) Flow coefficient

with V in m3/s, u2 in m/s and d2 in m.Flow coefficient j reflects the volumeflow discharged by a fan at a givenouter impeller diameter and blade tipspeed.

d) Power coefficient

λ is a measure of the shaft power re-quirement.

e) Diameter coefficient

This variable indicates by how manytimes the outside impeller diameterexceeds that of a reference fan with ψ= 1 and ϕ = 1.

f) Tip speed ratio

This parameter indicates by howmuch the impeller runs faster or slo-wer than the reference fan having ψ =ϕ = 1.

g) Throttle coefficient

τ is the parameter for the parabolicsystem graph in the dimensionlessfield of characteristic curves.

� = �pt · V̇

Pw

� = �pt · f*

· u22

2

� = V̇

u2 · � · d22

4

=� · �

�

=�

�

14

12

� =�

�

12

34

� =�2

�

2) Density and temperature changes

In ventilation and air-conditioning en-gineering, characteristic fan curvesare shown for a temperature of 20°C= 293 K. Density � is 1,20 kg/m3 atthis temperature. Where differenttemperatures apply (e.g. an outdoorfan to be rated for -15°C = 258 K) thefan’s characteristic curves for thattemperature can be obtained by con-version.

a) Volume flow always remains con-stant, i.e. a fan delivers the samevolume per unit time regardless ofwhether the air is „light“ (e.g.,+40°C) or „heavy“ (e.g., -15°C).This is because of the density (un-like the mass flow, which doeschange with temperature) notbeing a factor in the volumetric flowrate.

b) Values depending on density � andhence, on the temperature, willchange with it (refer to section 2.1).

Hence:

Total pressure increase ��pt, dyna-mic pressure ��pd, system resi-stance ��pt and power input requi-rement pW are all affected by thechange. The magnitude of theirchange is proportional to the chan-ge in density .

Summing up, we can write

This relationship applies to total pres-sure increase generated by the fan aswell as to system resistance.

��

Change in density (from �1 to �2, i.e. from +20°C to +15°C in this example)

2

1

·V1 = ·V2

�pt2 = �pt1 · = �pt1 · T1

T2

2

1�pd2 = �pd1 · = �pd1 ·

T1

T2

2

1Pw2 = Pw1 · = Pw1 · T1

T2

linear log-log

��

��

��

�

�

*)Neglected in ventilation and air-condition technics (�pt < 2500 Pa)

2


Comparison between RV/RA andAXN fans:

Flow coefficient:The RV fan has by far the highestflow coefficient (max. 1,2) whencompared to AXN (0,38) and RATR(0,55).

Pressure coefficient:RA fans have a steeper characteri-stic curve. This becomes evident ifwe compare deviations of the sy-stem characteristic curve A whichintersects the fan’s characteristiccurve at B. If the system characteri-stic curve A is lower than calculated(A1, point of intersection B1) or hig-her than calculated (A1, B1) in prac-tical system operation, changes inpressure coefficient and hence, vo-lume flow rates, remain small.

The situation is similar with the AXNfan, but it should be noted here thatstalling will occur from a certain flowcoefficient threshold onwards (inthis case, 0,23), i.e. an appropriateairflow over the blade profile is nolonger ensured.

Axial-flow fans must never be ope-rated in the stall range. They mustalways be dimensioned with an ap-

5.3 Selection criteria

Using the above dimensionless para-meters, it is now possible to comparethe main fan designs:

I. Backward curved blades �refer tosection 4.2.1�. (high-performancefan, abbreviated to „RA“)

II. Backward inclined straight blades�refer to section 4.2.2�. (high-per-formance dust fan, abbreviated to„RA St“)

III. Radial ending blades �refer to sec-tion 4.2.3�. (also referred to asconveyor fan, abbreviated to „RATR“)

IV.Forward curved blades �refer tosection 4.2.4�. (also referred to asdrum rotor-fan, abbreviated to„RV“)

All centrifugal fans are assumed topossess a spiral casing. Plug-infans are not taken into account inthe present selection criteria.

V. Axial flow fan with outlet guidevanes �refer to 3.2 and 3.3�. (abbre-viated to „AXN“)

Flow coefficient �

Pre

ssur

e co

effic

ient

� →

Pow

er c

oeffi

cien

t →

0 0,05 0,1 0,15 0,2 0,25 0,3 0,35 0,4

0,20 1,2

0,18 1,0

0,16 0,8

0,14 0,6

0,12 0,4

� = 0,620,67

0,73

B2

A2

A1

B1

A

B

0,72

0,68

0,84

�

0,79

0,82

0,10 0,2

→

Centrifugal fan with backward curved blades „RA“

Pow

er c

oeffi

cien

t →


0 0,05 0,10 0,15 0,20 0,25 0,30 0,35 0,40 0,45 0,50 0,55

0,50

0,45

0,40

0,35

0,30

0,25

0,20

1,4

1,2

1,0

0,8

0,6

0,4

0,2

0,59 0,70 0,750,79

0,800,79

0,780,72

0,64

� =

�

A2

A

A1B2

B

B1

→

Centrifugal fan with backward inclined straight blades „RAST“

Pre

ssur

e co

effic

ient

� →

0 0,05 0,10 0,15 0,20 0,25 0,30 0,35 0,40 0,45 0,50 0,55 0,60

0,80

0,70

0,60

0,50

0,40

0,30

0,20

0,10

1,4

1,2

1,0

0,8

0,6

0,4

0,2

A2 A A1

B2 BB1

0,42 0,58 0,670,75 0,76 0,77

0,760,74 0,72

0,710,69 0,68

� =

�

Flow coefficient � →

1,6

Centrifugal fan with radially ending blades „RATR“

Pre

ssur

e co

effic

ient

� →

Pow

er c

oeffi

cien

t →


2

Centrifugal fan with forward curved blades „RV“propriate safety margin separatingthem from the critical point.

RV fans have a flat characteristiccurve, i.e. slight pressure variationswill result in major volume flowchanges.

Efficiency:The RA fan has the highest efficien-cy (0,84), followed by the AXN unit(0,82). With a view to safety, only0,78 of this should be utilized on anAXN fan. RV fans, on the otherhand, achieve modest efficienciesat best (max. 0,69).

Power coefficientThe RA fan draws maximum powerat approximately its highest effi-ciency and should be dimensionedwith this characteristic in mind. Itmakes the fan safe against overloa-ding, given that the power demandwill decrease both when it is thrott-led and when volume flow increa-ses. Shaft power requirement of AXfans tends to be quite constant overthe rating range. RV units, on theother hand, exhibit a rapid increasein power demand when the volumeflow rises; an overload risk wouldtherefore exist if, e.g. the system re-sistance should turn out to be lessthan projected in theory.

Diameter coefficientThis is lowest in the case of theAXN fan (1,6 at ηmax), attesting tothis fan’s main advantage, viz. com-pact build. RV comes next at 1,8,followed by RA fans at 2,0.

Tip speed ratioThe highest ϕ and ψ values at mini-mum blade tip velocities are achie-ved by the RV fan (σ = 0,36), com-pared with 0,6 on RA and 0,95 onAXN units.

Axial-flow fan with outlet guide vanes „AXN“

Flow coefficient �0 0,2 0,4 0,6 0,8 1,0 1,2 1,4

6

5

4

3

2

3

2,5

2

1,5

1

� = 0,550,67

0,69 B2

A2

A1

B1

A

B0,5

0,35�

0,620,68

→

Pre

ssur

e co

effic

ient

� →

Pow

er c

oeffi

cien

t →


0,2 0,25 0,3 0,35 0,4

� = 0,81 0,82

B2

A2

A1

B1

A

B0,73

0,56

�

0,81

0,16

0,15

0,14

0,13

0,5

0,4

0,3

0,2

stall

→

Pre

ssur

e co

effic

ient

� →

Pow

er c

oeffi

cien

t →

2


5.4 Parallel operation

Where a very high volume flow is spe-cified, it is possible to operate two ormore fans in parallel. Double-inletcentrifugal fans are an example ofparallel operation, although here thetwo fans are rigidly interconnected. Ina classic parallel fan arrangement theindividual units are run independentlyof each other. From a flow controlpoint of view, such set-ups are usefulfor increasing or decreasing through-put by bringing the fans selectively onstream.

To determine the characteristic curveof parallel fans it is necessary to addtheir volume flows at identical �pt va-lues (as in the example of the RA11.1, NG 800 centrifugal fan).

V1 = Characteristic curve of one fan

V2 = Joint characteristic curve ofboth fans

B1 with .V1 and �pt1 =

operating point when one fan is run-ning.


operating point when both fans arerunning.

5.5 In-line/series operation

To overcome exceptionally high resi-stances, two or more fans may be ar-ranged in series. In this configurationtotal pressures �pt would theoretical-ly have to be added, while

.V would re-

main constant. However, this is notachievable in practice. A real-life sy-stem of this type encounters losses,chiefly due to inferior inflow conditi-ons prevailing at the second stage.

V1 = Characteristic curve of one fan

V2 = Joint characteristic curve ofboth fans


operating point when one fan is run-ning.


operating point when both fans arerunning.


2

5.6 Pressure measurement on fans

In aerodynamic engineering it is stan-dard practice to treat pressures abo-ve the atmospheric pressure po (baro-meter reading) as absolute values.This is acceptable if the ambient airpressure is taken as the „zero“ refe-rence level. As a result, one may ob-tain negative static pressures, for in-stance on the inlet side of fans.

The total pressure difference across afan is the difference between totalpressures on its inlet and outlet side

�pt = pt2 - pt1 = ps2 + pd2 - (ps1 + pd1)

= ps2 - ps1 + pd2 - pd1

= �ps + �pd

In other words, the total pressure dif-ference is the sum of the static pres-sure difference �ps and the dynamicpressure difference �pd between thefan inlet and outlet side (with �ps and�pd being measured as mean valuesacross the fan’s inlet or outlet cross-section, respectively).

Examples of measuring arrangements on centrifugal fans

a) Outlet side resistances, free fan inlet

�pt = ps2 + pd2 = pt2

= ps2 + c22, da pt1 = 0!

�2

b) Inlet side resistances, free fan outlet

�pt = ps1 - pd1 + pd2

For the particular case that A1 = A2 we obtain�pd1 = pd2

Hence pt = ps1

1. without diffuser

2


�pt = ps1 + pd3 - pd1

2. with diffuser

c) Outlet and inlet side resistances

�pt = ps2 + ps1 + pd2 - pd1

For the particular case that A1 = A2 we obtainpd1 = pd2

Hence �pt = ps2 + ps1.


2

d) Measuring arrangement for a cen-trifugal fan

While the mean dynamic pressurecan be obtained from the measuredvolume flow, static pressure is moredifficult to determine, particularly onthe fan outlet side, and there exist se-veral options for doing so. Characteri-stic curve data should therefore beaccompanied by a description of themeasuring set-up employed. Thus, itis important to know in the case of acentrifugal fan whether the staticpressure was determined on its inletor outlet side, and in the latter case,one should indicate at which pointdownstream of the fan the measure-ment was taken (i.e. directly behindthe guide vanes or at some distancefrom them).

In the present example pressuremeasurement is taken on the inlet si-de, with a screen restrictor simulatingthe upstream system resistance. Inpreparing the characteristic curves,the dynamic pressure over the entirecross-section is then added arithmeti-cally to the static pressure reading.Measurements with and without out-let side ducting give the same results.On the other hand, static pressure va-lues determined on the outlet side willvary according to whether the probeis mounted directly downstream ofthe guide vanes or at some distanceinto the ducting. This is due to the ringcurrent exiting the guide vane assem-bly; a certain flow path is necessaryfor the medium to become homoge-neously distributed again. As it doesso, part of the dynamic pressure isconverted into static pressure (pres-sure recovery), while the remainder islost as so-called hub impact loss.

Example:On an axial-flow fan with a hub ratioof 0,56, the mean dynamic pressurein the ring flow is equal to:

Velocity profiles: (1) upstream of the axial-flow fan(2) directly downstream of guide vanes(3) 2-4D downstream of fan

V̇ = c1 · A1 = cR · AR

cR = 1,457 · c1 bzw.

pdR = 1,4572 · pd1 = 2,12 · pd1

c1 · = cR · d12 - (0,56 d1)2 = CR · d1

2 · (1- 0,56)2

= CR · d12

· 0,6864

d12 · �4

�4�4

�4 � �· ·

·

�pCalibratedmeasuring

nozzle

�ps1Adjustablescreen restric-tor

fan with outletguide vanes

without outletguide vanes

without outlet guide vanes, with diffuser

2


From the above it is evident that thedynamic pressure in the ring flow ismore than twice as high as the pres-sure measured across the entire ductarea.

The hub loss, according to 2.4.2.1,can be written as

�p = (cR - c3)2 = 0,21 pd3 = 0,21 pd1

This must be viewed as the „inherentloss“ of each fan. It is already ac-counted for in the characteristic curveif the measurements are taken a suf-ficient distance downstream of the im-peller.

If the measurements are taken on theinlet side on a fan without downstre-am ducting and the dynamic pressureof the ring flow is included in the totalpressure difference, then the hub lossis not reflected in the characteristiccurve. This fact would need to be ta-ken into account at the fan dimensio-ning process.

e) General notes

If it is intended to take static pressuremeasurements via orifices in the ductwall, several orifices should bespread evenly over the duct circumfe-rence. These should then be inter-connected via a ring line. This is thebest way to compensate for variationsand to obtain a mean value. The sta-

tic pressure can only be assumed tobe near-constant over the cross-sec-tional area if the flow lines at the mea-suring point are straight. This will notbe the case downstream of elbows(refer to section 2.7), fittings and baf-fles. If conditions are not right for apressure measurement via wall orifi-ces, the cross-sectional area must bescanned with a pressure probe, andthe mean value must be determinedfrom the grid point readings.

Acceptance and performance mea-surements are governed by VDI Gui-deline 2004, which describes all de-tails concerning test set-up and exe-cution.

2

VI. Speed control

In the following paragraphs, theterm „fan control“ is deemed to re-fer to the control of the volumeflow.

6.1 Throttle control

The most straightforward but least ef-ficient control method is that of thrott-ling the flow. An adjustable restrictingdevice is fitted into the system to varythe system’s characteristic curve.The position of the points of intersec-tion with the fan curve will thus bechanged, i.e. shifted to the left (smal-ler volume flow).

As an example, let us again considerTLT Turbo GmbH’s RA 11.1 / NG 800fan.

Efficiencies at the points of intersec-tion:

B : 83 % B1: 84 % B2: 82 %

B3: 77 % B4: 70 % B5: 63 %

It is evident from this example thatdue to throttling of the flow, the cha-racteristic curve of the fan is intersec-ted further to the left, i.e. at a higherpressure, which requires additionalthrottling. Moreover, efficiency of thefan is reduced as the degree of thrott-ling increases.

Throttling the volume flow V by about25% from its level at intersection pointB will bring down shaft power inputdemand from PW = 21,1 kW to PW’ =20,2 kW. This is equivalent to a 4%decrease.

V̇ in m3/s

�


2

6.2 Blade pitch control

A more efficient, but also more com-plex method is that of controlling thefan speed (rpm) via appropriate varia-ble-speed electric motors. One ad-vantage of this control approach isthat the fan can always be operated ina favourable efficiency range. Thecharacteristic curve of the system re-mains the same, while that of the fanwill change according to the propor-tionality laws. On the downside, thiscontrol method involves higher capi-tal outlay due to the cost of the elec-tric frequency converter, as well as in-ferior efficiencies in part-load operati-on.

Example:

Efficiency is 83% at all points of in-tersection!

Reducing the volume flow V by about25% from its level at intersection pointB (i.e. to a value corresponding ap-proximately to intersection point B3)will bring down shaft power input de-mand from PW = 21,1 kW to PW’ = 8,8kW.

This is equivalent to a 58% decrease.The gain achieved over mere thrott-ling is obvious.

6.3 Blade pitch adjustment

On axial-flow fans with blade pitch ad-justment, volume flow can be control-led by changing the blade angle.

Example:


B : 77 % Bo: 78 % B1: 70 %

B2: 59 % B3: 50 % B4: 40 %

B5: 30 %

Reducing the volume flow V by about25% from its level at intersection pointB (i.e. to a value corresponding ap-proximately to intersection point B2)will bring down shaft power input de-mand from PW = 10,4 kW to PW’ =5,7 kW. This corresponds to a 45%decrease.

Controlling the volume flow of axialfans via the blade pitch setting will notquite yield the efficiencies achievedby rpm control. On the other hand, theassociated electrical losses are elimi-nated.

The investment cost of an axial-flowfan with „blade pitch adjustment un-der load“ (i.e. on the moving fan) issignificantly higher than that of anequivalent unit whose blades can on-ly be adjusted when stationary. Theadditional expense will generally payoff only if, in the specific operating en-vironment, blades must be adjustedvery often.

6.4 Inlet vane control

Flow control vanes can be fitted onthe inlet side of both centrifugal andaxial-flow fans. Acting as adjustableguide vanes, they modify the directionof the inlet velocity c1 into the impel-ler. By imparting an angular momen-tum (swirl) at the impeller inlet, theyproduce a change in volume flow.

Example: (TLT-Turbo GmbH centrifu-gal fan, RA 11.1, NG 800, with adju-stable inlet vanes)


B : 83 % B1: 80 % B2: 60 %

B3: 40 % B4: 30 %

Reducing the volume flow V by about25% from its level at intersection pointB (i.e. to a value corresponding ap-proximately to intersection point B2)will bring down shaft power input de-mand from PW = 21,1 kW to PW’ =12,5 kW. This corresponds to a 41%decrease.

For large volume flow changes, inletvane control makes sense - due to thesteep efficiency decline - only whencombined with a pole-changing motor.For instance, a pole-changing motorwith three speeds (100, 75 and 50% ofnominal rpm) offers a broad controlrange at an optimum efficiency.

Benefits of flow control based on ad-justable inlet vanes include low in-vestment cost and the fact that squir-rel cage motors may be used.

2


VII. Drive unit dimensioning

7.1 Motors

Power demand PW on the fan shaftcan be calculated (refer to section5.1). It is common practice to add cer-tain power reserve to the calculatedrequirement PW. The amount of thismargin is typically 5-10% for direct-drive fans and 10-20% on belt-drivenunits, depending on the size.

An important motor selection para-meter is its accelerating torque. Thismust be in a certain proportion to thefan’s moment of inertia, the fan thusstarting up properly.

The mass moment of inertia J refersonly to rotary fan components, i.e.impeller, hub and shaft. It is the pro-duct of multiplying the mass of theserotating parts with the square of theso-called „inertia radius“. Typically,this parameter is determined experi-mentally and stated by the fan manu-facturer. Motor manufacturers usuallyaccept an acceleration time of 10 se-conds. The selection of the motor canthus be validated using the followingexpression:

tA = acceleration time in seconds

J = mass moment of inertia of thefan wheel and motor in kgm2

nM = motor rpm

Mb = mean acceleration torque inNm, calculated as the differen-ce between motor torque Mm

and fan torque Mw

The above equation applies to direct-driven fans. In the case of V-belt drivesystems, the so-called reduced mo-ment of inertia must be used:

Jred. = JM + JV

JV = Impeller moment of inertia

JM = Motor moment of inertia

Jred. = Sum of the moments of inertia(JV + JM)

The torque MW can be calculatedfrom shaft power PW and the fanspeed nv. The accelerating torque Mbcan be obtained from the motor ma-nufacturer.

7.2 V-belt drive

V-belt drives are widespread in venti-lation and air-conditioning equipment.A V-belt has a very good adhesion,being ‘wedged’ into the groove of thepulley. It should be dimensioned soas to ensure that the belt speed willnot exceed 20 m/s. Belts are selec-ted in accordance with DIN 2218 onthe basis of manufacturers’ cataloguedata, which allow the engineer to de-termine a given V-belt’s power trans-mission capability as a function of beltprofile, pulley diameter and rpm

7.3. Couplings

Couplings serve to connect rotarymachine components - in the presentcase, they link the motor to the fan.They are required to transmit a torqueM at a given rotational speed.

As a result, the main coupling dimen-sioning parameters are fan speed nvand fan shaft torque MW, or shaft po-wer PW, respectively. The correlationcan be written thus:

Mw = fan torque in Nm

Pw = shaft power in kW

nv = fan speed in rpm

The couplings used in ventilationand air-conditioning applications aretypically of the resilient, direct-actingtype. In special cases - e.g., if themotor does not attain its nominalr.p.m. within the maximum accelera-tion time - it is possible to use centri-fugal clutches. They allow the motorto run up to its nominal speed first,while the fan is then accelerated toits operating r.p.m. via friction forceswith an appropriate time lag.

tA =

J · �Mb

� · n30

J ·nM

9,55 · Mb

Jred ·nM

9,55 · Mb

nv

nM( )2

�·n30

Pw

�

tA =

Mw = bzw. mit � =

where: � = ; tA =

whereMw = 9549 ·Pw

nv


2

VIII. Explosion protectionon fans (current status Jan. 2005)

8.1 Standards situation

Since the enactment of ATEX 100,previous national regulations such asVDMA standard sheet 24169, Parts 1and 2, are no longer applicable.

Although the relevant European pro-duct standard for fans is still in thedraft phase, Parts 1-7 of DIN EN13463 already exist.

EU Directive 94/9/EC (ATEX 95) re-gulates the approximation of the lawsof European Union member statesconcerning equipment and protectivesystems intended for use in potential-ly explosive atmospheres.

ATEX 137, or Directive 1999/92/EC,stipulates minimum regulations forthe safety and health protection ofworkers at risk from potentially explo-sive atmospheres.

While ATEX 95 addresses manufac-turers of equipment, components andprotective apparatus, ATEX 137 co-vers the installation of equipment andadaptations of existing systems.

The above directives have been ap-plicable in Germany since July 1,2003.

Basic requirements on the design,construction, testing and marking ofnon-electrical equipment are definedin the European standards seriesprEN 13463, Parts 1 - 8.

Fans in a general sense are treatedas non-electrical equipment in this se-ries, which contains the following spe-cific standards:

DIN EN 13463-1, April 2002: Non-el-ectrical equipment for potentially ex-plosive atmospheres - Basic methodand requirements, with amendmentsof July 2003

pr EN 13463-2: Non-electrical equip-ment for use in potentially explosiveatmospheres - Protection by flow re-stricting enclosure

pr EN 13463-3: Non-electrical equip-ment for use in potentially explosiveatmospheres - Protection by flame-proof enclosure

pr EN 13463-4: Non-electrical equip-ment for use in potentially explosive

atmospheres: Protection by inherentsafety

EN 13463-5, March 2004: Non-elec-trical equipment for use in potentiallyexplosive atmospheres - Protectionby constructional safety

pr EN 13463-6: Non-electrical equip-ment for use in potentially explosiveatmospheres - Protection by controlof ignition source

pr EN 13463-7: Non-electrical equip-ment for use in potentially explosiveatmospheres - Protection by pressu-rization

pr EN 13463-8: January 2004: Non-electrical equipment for potentiallyexplosive atmospheres - Protectionby liquid immersion

EN 50303, Group 1, category M1equipment intended to remain func-tional in atmospheres endangered byfiredamp and/or coal dust

DIN EN 1127-1, Oct. 1997: Explosiveatmospheres - Explosion preventionand protection - Part 1: Basic con-cepts and methodology

DIN EN 1127-2, July 2002: Explosiveatmospheres - Explosion preventionand protection - Part 2: Basic con-cepts and methodology for mining

Other German standards include thefollowing:

DIN 14428, Sept. 1988: Firefightingequipment - Explosion-proof portabletransfer pump with electric motor -Requirements, type and acceptancetest

DIN 14427, March 1995: Firefightingequipment - Explosion-proof portabletransfer pump for dangerous fluids,with electric motor - Requirements,testing

DIN 14642, Oct. 1995: Portable sear-chlight with mounting equipment forvehicles, explosion-proof

DIN 22419-1, Nov. 1995: Electricalapparatus for potentially explosive at-mospheres for mining - Cable entries- Part 1: Safety requirements andtesting

DIN 22419-2, Nov. 1995: Electricalapparatus for potentially explosive at-mospheres for mining - Cable entries- Part 2: Gland adaptors for entries;safety requirements and testing

DIN 22419-3, Nov. 1995: Electricalapparatus for potentially explosive at-mospheres for mining - Cable entries- Part 3: Gland flanges for entries; sa-fety requirements and testing

DIN EN 50016 (VDE 0170/0171, Part3), May 1996: Electrical apparatus forpotentially explosive atmospheres -Pressurized apparatus "p", Germanversion, EN 50016:1995

DIN EN 50039, April 1982: Electricalapparatus for potentially explosive at-mospheres - Intrinsically safe electri-cal systems "i" - (VDI specification forelectrical apparatus for potentially ex-plosive atmospheres for mining)

DIN EN 50050, June 2002: Electricalapparatus for potentially explosive at-mospheres - Electrostatic hand-heldspraying equipment; German versionEN 50050:2001

DIN EN 60079-10 (VDE 1065 Part101), Sept. 1996: Electrical appara-tus for explosive gas atmospheres -Part 10: Classification of hazardousareas (IEC 60079-10:1995) Germanversion EN 60079-10:1996

DIN EN 60079-14 (VDE 0165 Part 1),Aug. 1998: Electrical apparatus forexplosive gas atmospheres - Part 14:Electrical installations in hazardousareas (other than mines) (IEC 60079-14:1996); German version EN 60079-14:1997)

DIN EN ISO10807, Jan. 1997: Pipe-work - Corrugated flexible metallic ho-se assemblies for the protection of el-ectrical cables in explosive atmos-pheres (ISO 10807:1994); Germanversion EN ISO 10807:1996

DIN VDE 0170/0171-9, July 1988 El-ectrical apparatus for explosive gasatmospheres; protective encapsulati-on "m" German version EN50028:1987

DIN VDE 0170/171-13, Nov. 1986:Electrical apparatus for potentially ex-plosive atmospheres; requirementsfor apparatus in zone 10

DIN VDE 0848-5, January 01: Safetyin electric, magnetic and electroma-gnetic fields - Part 5: Protectionagainst explosion

2


It is recommended to track the har-monization of standards and theirtransposition into the national sy-stems in the EU's Official Journal andin the German Federal Gazette (Bun-desanzeiger), e.g., at

http://europa.eu.int/comm/enterpri-se/nando-is/cpd

and

http://bundesanzeiger.de

8.2 Product standard for fans

The European product standard forfans is available in draft form as pr-DIN EN 14986, June 2004.

The title of this draft standard is De-sign of fans working in potentially ex-plosive atmospheres.

Compared to the national code (VD-MA standard sheet 24169, Parts 1and 2), this document imposes anumber of changes.

Thus, the following information mustappear on the nameplate:

Apparatus group: I or II; a distinctionis made according to whether theequipment is intended for use in mi-ning or other applications.

Apparatus category: Categories 1through 3 express the requisite level

of equipment safety to be met by themanufacturer through appropriate de-sign

Conveyed medium: G = gas, D =dust, or GD = gas/dust mixtures

Ignition protection type: Indicatesthe design safety of apparatus andequipment, with requirements on ma-terial combinations, gap dimensions,V-belt, anti-friction bearings, etc.

Explosion group: Defines the typeof potentially explosive gas atmos-phere in which the equipment is to beused

Temperature class: Defines the ac-ceptable maximum surface tempera-ture on the apparatus

8.3 Marking example:

MarkApparatus

groupApparatuscategory

Conveyedmedium

Ignitionprotection

Explosiongroup

Temperatureclass

Marking ofapparatusfor use inpotentiallyexploxiveatmospheres

I: Mining

II: All otherapplications

for II:

1: even in caseof rateequipmentmalfunctions

2: even in caseof frequentequipmentmalfunctions

3: in normaloperation

for II:

G: Gases,vapours,fumes

D: Air/dustmixture

GD: Gas and dust

Materialcombinations

Gapdimensions

V-belts

Anti-frictionbearings

etc.

II AII BII C

T1: max. 450° CT2: max. 300° CT3: max. 200° CT4: max. 135° CT5: max. 100° CT6: max. 85° C


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8.5 Explosion protection of fans, il-lustrated for a direct-driven centri-fugal fan

The design of fans for use in potenti-ally explosive atmospheres typicallyinvolves the following steps:

� An increased gap between theimpeller and the inlet nozzle, plusthe selection of appropriatelymatched impeller/nozzle materi-als.

� Use of non-contacting (labyrinth-type) shaft seals to prevent heatbuild-up, with prevention of air le-aks via an additional bypass lineto the inlet. The impeller can beprovided with reverse flow bladesfor pressure relief.

� Installation of long-life antifrictionbearings, special hub-to-shaftattachments to prevent shifting,and fitted pins for securing thebearing housing.

� Use of driveshafts of appropriateflexural strength, separation ofcritical and operating rpm by anappropriately wide margin.

Electrostatic discharge devices(refer to the grounding sketchacross)

5. Two carbon brushes are spring-biased against the fan input shaft.Any electrostatic charge will thusbe neutralized via the carbon brus-hes, their brass holder, and a cu-stomer-supplied grounding electro-de.

On the process side, the custo-mer should ensure that the fanscannot suck in any foreign matterwhich might deform componentsor generate sparks.

� Hazardous-duty fans should al-ways be directly driven via a cou-pling.

Impeller with reverse blades

�

�

Driveshaft

Copper

Grounding electrode

Foundation

�

� �

�

�

8.4 Design notes

Some design recommendations ex-tracted from the product standard aregiven below.

Category 1 : Gas

• All category 2 requirements must bemet.

• Taper-lock hubs and V-belt drivesare not permitted

• Tests for gas-tightness must beconducted

• Flame inhibitors must be fitted onthe inlet and outlet connection

• Category 1 units for outdoor usemust conform to prEN 13463-3

Category 2 : Gas and dust

• All category 3 requirements must bemet.

• Units with power inputs exceeding5.5 kW must not have taperlockhubs.

• Fan housings must be continuouslywelded

• Anti-friction bearings to be rated fora minimum service life of 40,000hours

Category 3 : Gas and dust

• Protection against ingress of foreignmatter

• Accretion

• Fan drive and coupling per DIN EN13463-5

• Shaft seal, anti-friction bearings,brakes and brake systems mustconform to DIN EN 13463-5

• Units with power inputs exceeding15 kW must not have taperlock hubs

As regards impeller-housing materialcombinations, it is recommended towait for the final vote of the prEN14968 product standard.

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9.1 Free-inlet fans without inletnozzle

The characteristic curves of a fan arealways measured with an inlet nozzleon the manufacturer’s test rig. If thenozzle is eliminated, as shown above,the flow lines will show the illustratedpattern due to the sharp flange edges.Restriction of the flow occurs, resultingin unfavourable impeller inlet conditi-ons. These in turn give rise to perfor-mance losses, i.e. the fan fails to attainthe rig-testing characteristic curve sta-ted in the performance data.

9.2 Free-outlet axial fan

Let us return to the example from sec-tion 5.6, paragraph d), where it was cal-culated that a hub ratio of 0.56 would gi-ve the outlet velocity CR = 1.46 c1 andthe dynamic pressure pdR = 2.12 pd1.

With this mounting configuration thepressure recovery will be lost. Accor-ding to section 5.6, it is equal to 2.12pd1 - 1.12 pd1 = 1.0 pd1, given that thehub impact loss amounts to 1.12 pd1(here it must be checked if this impactloss is accounted for by the characteri-stic curve, given the measuring methodemployed).

The loss of the pressure recovery of 1 xpd (related to the full cross-sectionalarea of the duct) must be added to theremaining system resistances whencomputing flow resistance.

It should also be noted when calcula-ting system resistances that any equip-ment fitted directly downstream of thefan (e.g. air heaters) will be exposed tohigher flow velocities in the blade ringarea, which in turn will result in higherdrag levels.

IX. Installation anddimensioning notes

In dimensioning a fan that is selectedon the basis of measured characteri-stic curves, care should be taken tocompare the envisaged field installa-tion scenario with the measuring set-up used in determining the curves.Fans are quite often fitted under unfa-

vourable flow conditions deviatinggreatly from the rating situation, re-sulting in them being unable to attaintheir operating point on the characte-ristic curve. The following notes areintended to address this circum-stance.


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The situation can be improved by in-stalling such equipment on the inletside, or by providing an appropriatediffusor.

In this case, all other factors beingequal, the outer diffuser diametershould be 1.25 times the diameter ofthe axial-flow fan.

The cross-section of the ring thus be-comes

The above gives CR3 = 0,8 c1 and pdR3

= 0,64 pd1. The outlet impact loss canthus be reduced significantly.

Diffusers are very sensitive in termsof their fluid dynamics since fan outletflow will never be quite regular. As aresult, the fluid flow may becomedetached from the diffusor wall. Suchstall effects increase the resistancecoefficient ζ.

Also, the diffusor and its losses mustactually be deemed part of the ductsystem, which makes it difficult to mo-del its behaviour. Preferably, fan anddiffuser should be measured as an in-tegral unit, as described in the mea-suring arrangements in section 5.6.

Another option for reducing outlet los-ses and improving the flow in down-stream elements lies in installing abaffle-plate diffuser, or radial diffuser.

Optimum values for

have been determined through expe-riments.

AR3 = –(1,25D)2�4

(0,56D)2�4

� 0,15 und � 1,5bD

D’D

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9.3 In-duct fans

When installing a fan into a duct sy-stem, care should be taken to minimi-ze interference and ensure maximumflow uniformity on the inlet and outletside. Arrangements in which the faninlet is located directly downstream ofsudden duct expansion or restrictionpoints, elbows, etc., should be avoi-ded. In particular, the inlet flow shouldnot come at an angle or with angularmomentum (swirl), since this maycause stalling in the impeller or othersevere forms of performance loss.


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9.4 Parallel and in-series operation

With parallel fan configurations, aproblem may arise if the characteri-stic curves of the individual units havea peak or turning point (as is verymuch the case with axial-flow fans).The resulting characteristic will thenshow the following pattern:

Since the loop in the resulting charac-teristic curve lies close to the apex, aconfiguration of this type may havethree operating points (1, 2, 3) bet-ween which the fan alternates (unsta-ble operation). When dimensioning afan for such a system, care must betaken to ensure that one operatingpoint will be located sufficiently far tothe right of the peak (which would be

the stall point in the case of axial-flowfans).

If centrifugal fans are arranged in se-ries for pressure boosting purposes,their inherent design will usually re-quire an extended length of ductingbetween the outlet connection of thefirst fan and the inlet of the secondone. This interconnecting duct canusually be provided with featureswhich ensure a proper inlet flow sup-ply to the second stage. In this case,the in-series configuration can be ex-pected to have a y value equal to thesum of its individual counterparts.

With axial-flow fans, the two stagesare typically mounted one directly be-hind the other. The disturbed outletflow from stage 1 will thus have an im-mediate effect on the inlet of stage 2.As a result, the pressure coefficientshould not be expected to exceed avalue of about 1,6.

B1: Operating point with one fan run-ning

B2: Operating point with both fansrunning

A1: System curve too high, instablerange

A2: Properly dimensioned system

10. Overview of old and new units of measurement

Conversion / Relationships

a) Force: 1 kp � 9,81 N = 9,81 ;1 N � 0,102 kp

b) Pressure: 1 mm WS � 1 kp/m2 �9,81 Pa � 0,0981 mbar

1 Pa � 0,102 mm WS �0,102 kp/m2 � 0,01 mbar

1mbar � 100 Pa � 10,2 mm WS� 10,2 kp/m2

1 torr = 1 mm Hg = 1,33322 mbar= 133,32 Pa

* Flywheel effect GD2 and mass mo-ment of inertia J are related by

GD2 = 4 g · J with g = 9,81 J in kgm2

G in ND in m

kgms2

ms2

SI unit Old technicalunit

Length

Time

Mass

Force

Torque

Energy

Specific gravity

Density

Velocity

Acceleration

Pressure

Frequency

Flywheel effect*

Moment of inertia*

Power

m m

s s

kgkps2

mkgms2N = kp

Nm kpm

Nm = J kpmN

m3( )kpm3

kgm3

kps2

m4

ms

ms

ms2

ms2

Nm2 = Pa kp

m2

s-1 = Hz s-1 = Hz

Nm2 kpm2

kgm2

Nms

= W kpms

, PS

�pt

or ψ

Individualcharacteristic

Resultingcharacteristic

.V or ψ

tlt english

Documents

backward curved

forward curved

correction

energy takes

elementary

elementary

outlet guide

systems characteristic