Fan Parameter

Overall shape and dimensions General

The Howden Cooling Fan delivery program consists of several product designs with fan diameters from 0.710 to 20 meter. Other geometric features are: 1. Blade number 2. Blade width 3. Blade shape (straight or swept forward) 4. Blade material (FRP or aluminium) Standard Howden Cooling Fans perform aerodynamic duties up-to 250 Pa fan static 3 pressure and 3000 m  /s air flow in wet and dry air-cooling installations.

Howden Cooling Fans impellers are designed for application in cooling towers, air-cooled condensers and air-cooled heat exchangers. The continues operating temperature range and allowable incidental upset temperature varies per product range, we therefor refer to the respective service manuals for the allowable temperatures per product type. The FRP impellers may be exposed to a higher maximum temperature for a short period of time, for example during start-up or stand still. In case the temperature of your installation exceeds the maximum allowable impeller temperature, then please contact Howden for review of the full operating conditions. In a humid arrangement FRP fan blades must be provided with an erosion-resistant layer on the inlet side in order to protect the blades from impact i mpact of water droplets (leading edge protection). The fan diameter, the fan rotation speed and the fan blade number are the particular parameters to match the performance of the fan to the aerodynamic duty point of the air-cooling installation. Blade width and blade shape are the principal instruments to reduce the noise generation of the fan. Product lines

Howden Cooling Fan product lines can be divided into four typical shapes as presented below. The blade pitch angles of all Howden Cooling Fans can be adjusted manually during standstill. For available diameters and number of blades see the diameter and blade number overview (PDF).

PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

ENF / KNF / ZNF fan

Classic straight aerofoil bladed standard fan with a normal noise performance and high efficiency.

ELF / KLF / ZLF fan

Standard low noise fan with straight aerofoil blades and high efficiency.

ELFA / ZVF fan

Very low noise fan with straight aerofoil blades which combines a good efficiency with a noise performance that exceeds the ELF / KLF / ZLF product lines.

SX fan

Sophisticated super low noise fan with a low number of forward swept blades. The SX program with its remarkable blade shape is the fan solution for total low noise projects.

Aerodynamic duty point The aerodynamic duty point of the fan is the combination of the Air Flow Q and the Fan Static Pressure FSP generated by the fan. The product of Q and FSP has the dimension of Power and is called the Aerodynamic

PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

ENF / KNF / ZNF fan

Classic straight aerofoil bladed standard fan with a normal noise performance and high efficiency.

ELF / KLF / ZLF fan

Standard low noise fan with straight aerofoil blades and high efficiency.

ELFA / ZVF fan

Very low noise fan with straight aerofoil blades which combines a good efficiency with a noise performance that exceeds the ELF / KLF / ZLF product lines.

SX fan

Sophisticated super low noise fan with a low number of forward swept blades. The SX program with its remarkable blade shape is the fan solution for total low noise projects.

Aerodynamic duty point The aerodynamic duty point of the fan is the combination of the Air Flow Q and the Fan Static Pressure FSP generated by the fan. The product of Q and FSP has the dimension of Power and is called the Aerodynamic

PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

Power Nair of the fan. Nair is the so called effective power output of the fan. The power input comes through the drive shaft and is named Drive Shaft Power Nsh. The ratio between Nair /Nsh is the static efficiency ηst of the fan.

Air flow The Air Flow is defined in [m3/s] at the temperature of the air when it passes through the fan. In principle the Air Flow has a value equal to the product of the average air velocity v in the flow section and the surface of that section. Since there is always a spread in the value of the air velocity over the section, for the determination of the average air speed the air speed must be read on several locations according to international standards. For instance the American Cooling Tower Institute (CTI) advises to do air velocity readings on at least 20 locations with a calibrated anemometer or pitot tube on equal flow sections in the fan inlet as close as possible to the fan. Due to r otation of the air, and flow "unfriendly" duct shapes it is hardly possible to do flow readings down stream of the fan. For the determination of its fan curves, Howden has built a test installation according to AMCA 210-74. Here the flow is measured over a calibrated nozzle. See figure 1. For air cooled installations this method is not possible due to the lacking of a nozzle. Mathematical relations

Pdyn = 0.5 * ρ * v2

{1}

Ptot = Pst + Pdyn

{2}

ηst = Nair /Nsh

{3}

1. Valve

5. Nozzle

2. Booster fan

6. Stream gauzes

3. Streamer

7. Engine frame

4. Stream gauzes 8. Test fan

Fig.1 Principal sketch of Howden aerodynamic test deviceaccording  AMCA210 

Static pressure PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

The Fan Static Pressure represents the flow resistance of an air cooled installation. It has the SI unit [Pa]. The value of the FSP is found according to the following definition:

FSP

= Pst2 - Ptot1 [Pa] = Pst2 - Pst1 - Pdyn1

{4}

See figure 2.

FSP = Pst2 - Ptot1 F1/F2 = 5 Fig. 2: Definition of FSP Perhaps the definition of the FSP is felt to be strange. However for an induced draught installation this definition of the FSP corresponds exactly with the static flow resistance of the heat exchanger section for which the influence of the velocity pressure is eliminated. This means that the value of Pst1 will differ from the FSP due to the influence of the velocity pressure. This can be better understand by deriving the theoretical value of Pst1 for different cases. The law of Bernoulli defines for an ideal flow without resistance the following relationship: Pst + Pdyn = Constant.

{5}

In any case in an flowduct f lowduct the theoretic value of the t he dynamic pressure is: Pdyn = 0.5 *ρ* v2

{6 }

For a configuration where no heat exchanger sectionis present andtaking the environmental environmen tal pressure zero (Pst2 = 0), according to {5}, Pst1= - 0.5 *ρ *ρ* v12

{7 }

See figure 3.

Whenthere isa heat exchanger section present with a flow resistance of RPa: {8} Pst1= - R - 0.5 *ρ * ρ* v12. Seefigure 4.

PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

Accordingto {4} The value of FSP is then: FSP= Pst2- Ptot1=Pst2- (Pst1+Pd n1), =0 - (- R - 0.5 *ρ* v12+ 0.5 *ρ* v12) = R Pa. Sothe absolute difference between FSP and P st1is: 2 ¦FSP¦-¦Pst1¦= -0.5 *ρ* v1

{9} {10}

ID/FD differences In principle there are two types of fan installations: a. Forced draught b. Induced draught For each type of installation the definition {4} (static pressure) must be interpreted as follows:

Induced Draught (ID) The heat exchanger section (flow resistance) is located up stream of the fan. For example, arrangement 5.5, 5.6, 5.7 and 5.8 in figure 5. In this case Pst2 = 0 if no diffuser is used. With diffuser Pst2 = ∆Pdiff. (See section 3-07.314, Pressure recovery by a diffusor) FSP is determined according to equation {9} and is equal to R. P st1 versus ambient pressure is negative and P dyn1 positive. Theoretically the value of Pdyn1 is equal to the reduction of P st1 according to Bernoulli's law {5} and corresponds with the difference between ¦Pst1¦ and FSP. According to {6} the value of Pdyn1 results from the air speed in the plenum. For an air speed v of 10 m/s this is 60 Pa. If the air speed is 2 m/s, Pdyn1 is 2.4 Pa.

Forced Draught (FD)

In this case the heat exchanger section (flow resistance)or principal flow resistance is down stream of the fan. For example arrangement 5.1, 5.2, 5.3 and 5.4 in figure 5.

The interpretation for FSP, in section 03-07.312 Fan static pressure, is made for an ID installation. For a FD installation the FSP is not equal to R but equal to P st2 since Ptot1 = 0. This follows from the definition:

2

FSP = Pst2 - Ptot1 = Pst2 = R - 0.5 * ρ * v2 - 0 In a FD Air Cooling installation there will be no pressure recovery from dynamic pressure to static pressure. Due to the presence of the bundles the additional kinetic energy will be

PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

dissapated. That is why an FD installation must be designed with a FSP which is at least equal to R.

For the identical heat exchanger section with a flow resistance of R in an ID configuration ¦Pst1¦ will differ from¦Pst2¦ in the FD configuration for two principal reasons: 1. Its theoretical difference: Pst1(ID) = - R - 0.5 * ρ * v1

2

2

Pst2(FD) = R - 0.5 * ρ * v2

2. Besides the velocity in main flow direction, there are rotational and turbulence components in the velocity which enlarge the real three dimensional value of v (v 2>v1). Consequently P st2(FD) is reduced. A contradictionary phenomenon is the idea that a rotating flow through a heat exchanger section has more resistance than a unidirectional flow. This due to the higher absolute speed of  the air and due toa less favorable direction of the air (more obstructions and longer air passage way) means an increa sing value of R.

In practice this means that the design of an FD air-cooling installation is more difficult to do by theoretical considerations only. Practical feed back from prototype tests will give the final information for adequate design

flow-related quality in general insensitivity to wind at fan inlet insensitivity to wind at fan outlet feasibility of guarantee measurements absence of annoying noise for operators suitability for high-temperature application life in case of wet application (erosion)

FD ID ± + -

+

+ -

± ±

-

+

+

± ±

+

PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

Diffusor When the air flows out the air cooling installation with a certain velocity v, according to the law of Bernoulli {3} (page 1, norm 03-07.311) it is possible to regain static pressure from the dynamic pressure and to reduce the FSP. However this works only for a fricton free flow. It means a uniform and swirl free flow. In reality it is only possible for an ID installation with an efficiency of 75 percent, and using a diffuser or fan stack with a cone angle of between 6 °-8.5°. (fig. 1). The pressure recovery ∆Pdif is calculated as follows: ∆Pdiff

= 0.75*0.5*ρ(vo2-vi2)

FSP need to be corrected with ∆Pdiff as follows for the case that a diffuser is used: FSPc = FSP + ∆Pdiff

{14}

PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

Flow disturbances Principles of flow disturbances

Besides the flow resistance of the heat exchanger there are other elements around the fan that have influence on the working of the fan namely: a. Fan inlet shape b. Flow obstacles c. Fan tip clearance a. Fan inlet shape

The performance of Howden standard fans are measured with an elliptic inlet bell with a length of 15 percent of the fan diameter and an elliptic ratio of 1:1.5. See figure 1. Other inlet shapes like: • • • •

inlet with radius cone flat-face flange cylindrical duct section,

have an unfavourable effect on the air flow around the airfoil of the fan blade. The inlet shapes will generate swirls and wakes which disturb the angle of attack of the air flow on the aerofoil of the fan blade. It is like the ingestion of turbulence by an air plane when it passes turbulence and the wing sections are flapping by the turbulence. See also figure 2.

b. Flow obstacles

In an air- cooling installation, heat exchanger sections and fan support structures are flow resistance elements but they also generate swirls and wakes. In an ID installation this will have the same negative effect on the air- flow around the blade aerofoils as the non -ideal bell inlet shape. That is why the influence of obstacles up stream of the fan is worse than down stream the fan. c. Tip clearance

The fan tip clearance has the following definition: cl = 2∗s/Df 

where :

{1

cl = tip clearance

PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

[mm]

s = gap between blade tip and fan ring

[mm]

The performance of Howden cooling fans are measured with a tip clearance of 0.01 (= 1%). A bigger tip clearance will have the effect of a leak; A smaller one has the effect of the closure of a leak, it means a higher pressure. In actual practice, the fan ring will never be truly round. The clearance 2s/Df  is the average value along the circumference. It is recommended to respect the following minimum local value: s min = 0.0025 Df 

{1

This minimum tip clearance value serves to prevent the blade tips from scuffling against the fan ring under changing operating conditions in the air cooling installation. (temperature increase, vibrations). The tip clearance has also influence on the fan efficiency. See section 03-07.325, power and efficiency.

Calculation of flow disturbance effects.

The interaction between disturbances and the fan is rela ted to the generation of swirls and wakes, which besides have a normal flow resistance effect, also have a disturbance effect on the flow angle of attack and on the flow around the blade aerofoil. In order to find the correct FSP, the flow disturbance influence must be elaborated by the use of  characteristic correction pressure terms ∆Pi for each type of disturbance: FSPc = FSP + Σ ∆Pi

{1

FSPc = Corrected FSP = Additional pressure drop by i. ∆Pi i = Disturbance: inl = inlet obi = obstacle at inlet obo = obstacle at outlet tpcl = tip clearance The correction terms ∆Pi of the obstacles and the inlet device have in principle the structure of a flow resistance term: ∆Pi = k i ∗0.5 ∗ ρ ∗ vi

2

{1

k i = flow resistance coefficient of i [-] vi = characteristic air speed [m/s]

The influence of the tip clearance is directly defined as a ratio Rtpcl of the correction term ∆Ptpcl and FSP: ∆Ptpcl = Rtpcl ∗ FSP The flow resistance coefficients for the inlet type, the flow obstacles and the values of  Rtpcl for the tip clearance can be found in respectively enclosure 1,2 and 3. The characteristic air speed for all is the air speed through the fan section in the main flow direction: 2 2 vf  = 4Q/(π(Df- – Dh ) {19}

PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

Dh = Fan hub diameter.

Figure 3: Influences of various inlet shapes 

PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

Figure 4: Flow resistance coefficient K obi  for obstacles at inlet 

PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

Figure 5: Flow resistance coefficient K obo  for obstacles at inlet 

PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

Figure 6: Rtpcl as a function of the fan tip clearance s 

Wind influence Since a cooling fan operates at one side in the open air at a relatively low pressure, wind surely influences the performance of the fan. For the same reasons as for the obstacles, the influence of the wind is more felt when it blows on the fan inlet side than when it blows on the wind outlet side. Special attention must be paid to situations where wind concentration effects arise by the air -cooling installation itself or by structures close to the air cooling installation. In particular vertical impellers with a horizontal shaft have an elevated sensitiveness for wind effects. However also horizontal fans installed at a great height and exposed to strong winds for sure are affected by winds. What happens is that a impeller comes partly or fully into stall which causes an elevated, PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

sometimes destructive dynamic stress level See figure: 1. Up to now the influence of the wind on the fan performance is only partly quantified. For instance the influence of the pitch angle is not clear. The same for the wind direction to the fan. Also fans beside each other have a mutual flow effect. A single fan "feels" the under or over pressure of its neighbor as an additional resistance. This phenomenon is amplified by wind. It can happen with strong winds that one fan performs perfectly and the fan beside is almost "dead". This "dead" fan is normally the fan on the up-wind side. All these aspects makes it clear that it is not easy to quantify the influence of the wind speed. The best suggestion is to take the dynamic wind pressure ∆Pwi as an additional static pressure drop. 2

∆Pwi = 0.5 * ρ * vwi

{20}

Fan scaling rules PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

Scaling of the fan operation points is done according to the so called fan laws: P :: ρ * Ut2

{21}

Q :: Ut * Df2 {22} In order to compare different fan configurations the fan duty point (Q,FSP) is transformed with help of the fan laws to the dimensionless figures C p and Cf. Cp and Cf have the following definition: {23} Cp = FSP/(0.5*ρ*Ut2) 2 {24} Cf = Q/(0.25*π*Df *Ut) Ut

= fan tip speed

[m/s]

{25} Ut = 0,5*ω*Df [rad/s] ω = angular rotation speed {26} ω = 2*π*RPM/60 By considering C p and Cf, the pure aerodynamic performance of the fan is considered without the influence of the: • • •

Fan diameter Fan rotation speed Air density (temperature)

By taking the dimensionless characteristics of the fan, the fan duty point can be compared with a model fan with the same shape, for instance a model fan in a test installation. By this way Howden is determining the characteristics of its fans. It has built a test facility according to AMCA 210-74. In this fac ility fan models with a diameter of  1829 mm (= 6') can be measured. By transforming the results into dimensionless figures, Cp,Cf ,ωst, the results can be applied to any fan diameter and rotation speed for fans with the same shape. The fan solidity σ the dimensionless figure which characterizes the aerodynamic effective shape of the fan. Its values is the total relative blade cord width or fan solidity σ. σ has the following definition: {27} σ = z*c/(π*Df)

Where: z

= number of blades

c

= blade width (cord)

It can be said that fans with the same σ and blade aerofoil, perform aerodynamically equally, i.e. for the same pitch angle they will always follow the same C p /Cf. This is the basic principle of the fan scaling rules and selection programs.

PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

Pressure margin About pressure margin and flow margin, Standard 661 of the American Petroleum Institute (API) defines the following: "Fan selection at design conditions shall ensure that at constant speed the fan can provide by an increase in blade angle a 10 percent increase in airflow and a corresponding pressure increase. Since this requirement is to prevent stall and inefficient operation of the fan, the resulting increased power requirement need not govern the driver rating." Supposing there is a square increase of flow resistance on a linear flow increase, the consequence of applying this (API) standard is that a 10 percent flow margin results in a 21 percent pressure margin. See figure 1. Another interesting value to know is which maximum FSP the fan can make for the set pitch angle before stalling and the corresponding Air Flow.

Surge limit PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

P1

D

ç

Allowable blade angle



é

Pressure

Resistance Lines è

Flow

Pressure margin = DP1/P1 * 100% or D P2/P2 * 100% Fig.1: Definition of Pressure margin according to API  661

Axial thrust For mechanical design features, it is interesting to know the value of the aerodynamic axial force on the fan. This force is called Axial Thrust and is calculated by multiplying the total pressure drop over the fan, P tot1,2 times the surface of the fan ring section. Fax = Ptot1,2 * ¼π * Dr2

{31}

where Dr = Diameter of the fan ring

Power and efficiency PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

The product of FSP and Q has the dimension of power and is the aerodynamic or effective power output of an air cooling installation. The power input to the installation comes through the fan drive shaft. This drive shaft power Nsh is the product of the shaft drive torque T sh and angular speed ωsh.

Power input:

Nsh = Tsh*ωsh

[kW] {28}

where: Nsh= Fan drive shaftpower Tsh = fan drive shaft torque ωsh

= angular rotation speed

[kW] [Nm] [Nm]

Effective Poweroutput Nair= FSP*Q

[kW]

{29}

where: As for every power transforming engine also for a cooling fan the ratio b etween effective power outp and power input is called the efficiency ηst. Remember {3}: ηst

= Nair /Nsh

For fans which are built in a ducting the FSP is not an interesting value. For that cases P tot1,2 is considered. Consequently there also exist an expression for the total efficiency: ηtot = Ptot2,1 *Q/Nsh.

Correction for deviating tip clearance

The performance of the Howden standard fan curves is measured with a tip clearance of 1 percent. Deviating values of the tip clearance result into deviating fan PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

{30}

efficiencies which need to be corrected as follows: shaft

1

= Nshaft

*

(1-KR)

The following correction must be made for calculation of the fan shaft power. Nshaft 1 = fan shaft power after correction for inlet shape, obstacles and diffuser  KR = Dη stat / η stat = correction factor for efficiency at deviating clearance Figure 1. has been plotted for some arbitrary 2s / Df values. For the applicable clearance design value or value measured in the installation - other the KR value can be determined by means of interpolation.

PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

Fig 1. Influence of tip clearance on fan efficiency 

PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

Field performance Introduction

It is a normal interest to verify the performance of a cooling fan in its installation. However this is not easy. The reason is that a cooling fan operates in unstable turbulent conditions, in both senses of th word. The main feature of a cooling fan is that it makes a big air flow over a relatively low pressure drop. Both parameters are hard to measure. A big air f low and a low pressure drop can only be mad when the cooling fan has a free air access and a free air outlet. This makes the air cooling installatio sensible for wind and other external disturbances. There will also be velocity variations over the various flow sections, which complicates the determination of air flows. Tests

In principle there are two possibilities to verify the performance of a cooling fan: 1. Scale model test A scale model test can be done with a geometric identical shaped fan on a well conditioned test facility. Example: the Howden 6’ test facility according to AMCA 210. Procedure: Howden 16-07.002 which is available on request. See also ISO 5801 2. Field performance test Like it has been explained in the introduction: Field performance tests are complicated. That is why it can only make sen to do it when international standards are carefully applied. Useful and practical is the Recommended Practice For Airflo Testing of CTI. Also DIN 24166 is very helpful. This last standard defines clearly the accuracy which can be expected. Power station and industrial applications in unstable environments like air cooling installations, are classified in category and 3. Those classes define the following accuracy’s:

Variable Class acc to DIN24166 Air Flow

2

3

+/-5%

+/-10%

Pressure drop Drive power Efficiency A-Sound Power Level

+/-5% +/-8% -5% +4 dB(A)

+/-10% +/-16% +6 dB(A)

Also see ISO 5802

Noise basics The noise phenomenon is not easy to understand. From physics point of view it is the vibration of air frequencies, which can be heard by human: 20-18000 Hz. The vibrations correspond with very small air pressure variations. Beside by the human ear also by microphone the (sound) pressure variations can be observed. PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

Like for alternating electrical current, for both the air particle velocity v (this is not the sound velocity) a the pressure variations p there has been also defined an effective value, it is the so called r.m.s val (root mean square). When the air particle velocity v and the sound pressure p is considered, always t effective value is meant. The particle velocity v is proportional to the pressure p: v ~ p. Noise is transmitted like longitudinal waves. Taking a spherical surface F with equal pressure p there can be found an expression with the dimension of power P. {1} P ~ p∗v∗F [W] Because v~p, it can be stated that: {2} P ~ p2∗F [W] -5 The range of audible pressures is very big: from 2.10 Pa for just audible to 200 Pa for the pain threshold. Howe the human feeling for noise is far from proportional to that scale. It is found to be useful to express the terms of equation {2} in the logarithm of the dimensionless ratios which is mention decibels. Doing this the following quantities are derived: The Sound Power Level PWL or LW: PWL = 10 lg (P/P0)

[dB] -12

with reference value P0 = 10

W and

The Sound Pressure Level SPL or Lp:

SPL

= 10 lg (p2 /p02) [dB]

With reference value p0 2.10 –5 Pa. It is understood to be correct by knowing that the reference value for particle velocity v which is proportional to p, is 5.10 –8 m/s.

Applying all consequently on equation {2} results in the following useful expression: PWL = SPL + 10 lg F [dB]

This expression is that useful because it gives a relationship between the sound power of a source (PWL) and the audible value (SPL) at a certain position with respect to the source. Moreover since it not possible to measure sound power the expression is also the way to determinate the PWL of a source: This is done by measuring a SPL on a control area F where the SPL is supposed to be equ like e sphere around the source or for big installations at 1 meter distance. This method is well defin in several international standards like ISO 1940/1. The use of octaves and A-weighting

The human ear has a different sensitiveness/awareness for the various sound frequencies. That is why mostly a noise value is filtered according to a logarithmic deviation into an octave bands. The variable human awareness is elaborated by a correction, a so-called A-weighting, for each octave band as follows: Octave [Hz] A-corr. [dB]

63 -26.2

125 -16.1

250 -8.6

500 -3.2

1k 0

2k 1.2

4k 1

8k -1.1

The not A-weighted spectrum is called the linear spectrum. From the A-weighted spetrum an A weighted total value can be found by the logarithmic addition of the different A-weighted octave valu as follows: PWL(A)=10 lg (10 0.1lg Lw(A)31.5 +100.1lgLw(A)63+........ 100.1lgLw(A)8k)

[dB(A)]

PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

An A weighted spectrum can be recognized by its unit: dB(A)

Fan noise The quantification of the noise generation by a cooling fan is in principle achieved by using general accepted standards like ISO 1680. Just measuring the noise of a fan is not enough criteria to accuratel predict the noise performance of a cooling fan. You must also know about t he influence of the operatin conditions and dimensions that effect the noise performance. Moreover, if noise production must be reduced, an even more sophisticated understanding of the noise generating mechanism is needed. Fo relatively slow running fan like the propeller cooling fan, there are a few characteristic noise generating flow phenomena [1]. 1. The so-called "rotor self noise". It is the turbulent and laminar vortex shedding at the blade rear sections and at the blade tip. 2. The ingestion of turbulence in the main air-flow. This turbulence is generated by the heat exchanger, fan supports or other upstream obstructions. The turbulence leads to random variati ons in angles of incidence at blade leading edges, causing fluctuating blade loads and surface pressures over a broad range of frequencies. 3. Besides the broad-band noise levels, sometimes there will be discrete peaks of sound pressure associated with the

blade passing frequency. This frequency is the product of the fan rotation frequency and the number of blades. The noise is caused by the pressure pulsation that is generated when a fan blade is passing a sharp and close disturban such as a support beam.

Figure. 2: Different noise generation fields for an axial flow fan according to [1] 

From a more simple and practical point of view can be stated that the noise intensity of a cooling fan is related to the quantit and intensity of flow-generated swirls. For the quantification of the noise intensity and in order to compare one cooling fan configuration with another, it is necessary to have a relationship between the noise intensity PWL and important design parameters like pressure drop p, flow Q, the fan tip speed U tip and the fan diameter Dfan. Through years of research and field measurements we have developed the following formula:

The characteristic value C represents the influence of the fan shape on the noise generating phenomena or as said before the intensity and quantity of swirls. From formula (1) it becomes clear that especially the tip speed Utip has a strong influence on the sound power lev The correction terms ∆dB are related to characteristic noise mechanism in an air cooling installation: The influen of obstructions and the influence of the flow inlet shape. The correction term for the inlet shape covers the additional noise by deviating from the ideal elliptic bell inlet shape.

PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

Typical C-values for Howden Cooling fans are: ENF ELF ELFA SX

37 35 33

27

dB[A] dB[A] dB[A]

ZNF ZLF ZVF

35.5 35 33

dB[A] dB[A] dB[A]

KNF KLF

37 dB[A] 35 dB[A]

dB[A]

Bottom value, function of several parameters like, tip speed, diameter and pitch angle) From the total PWL value of a fan, a linear spectrum is calculated by a correction table that varies for each fan type and octa band:

Reference [1]

S.E. Wright (1976), The Acoustic Spectrum of Axial Fans, Journal of Sound and Vibration, 45(2), 165-223

Sound pressure level For many projects it is required to calculate the sound pressure level on a certain position with resp to the fan. This standard provides some calculation methods for this purpose. The method works on the sound pressure level calculation for both areas and positions according to figure 1 and 2. Induced Draught installations

PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

Fig. 1: Induced draught configuration with SPL positions and areas  Please note that due to the turbulent airstream it is not possible to accurately predict the sound pressure level at 1 mete directly above the fan ring. Furthermore it is not possible to meassure the sound pressure level at 1 meter below the fan ring because of the presence of the air cooler. Positions:

A: 1 m beside B: 1 m above, 45° from the fan ring or the diffuser Areas

1 and 2 (0.5 Do + 1 < R < 5 Do) according to DIN 45635 P46) Formulas:

For A, B and area 2: SPL = PWL - 2 - 10 logF + Cspl 1 + Cspl 2 F = control area Cspl 1 = direction correction Cspl2 = near field correction F = 2pR2 Cspl1 = 2 - 6.8 (1 - Öcosa ) (0°£a£ 90°) If R£ Do, Cspl2 = 4(1-R/Do), else Cspl2 = 0 For area 1 SPL = PWL - 10 log (2pR(R + h)) h = height of the fan ring or diffuser from ground level Forced Draught Installations

PDF created with FinePrint pdfFactory Pro trial version http://www.fineprint.com

Fig. 2: Forced draught configuration with SPL positions and areas  Caution For a FD installation, reflections by ground surface can result into deviation of SPL levels which are out of t he scope of this consideration A: 1m beside

Positions A: 1m beside B: 1m below, 45 °from the edge of the inlet device Areas

1 and 2 (0.5Do +1< R