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 pressure and 3000 m3/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 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 aerodyna-mic 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).

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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



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 rotation 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



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] {4} = Pst2 - Pst1 - Pdyn1

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 the theoretic value of the dynamic pressure is: Pdyn = 0.5 *ρ* v2 {6 } For a configuration where no heat exchanger sectionis present andtaking the environmental pressure zero (Pst2 = 0), according to {5}, Pst1= - 0.5 *ρ* v1

2 {7 } See figure 3.

Whenthere isa heat exchanger section present with a flow resistance of RPa: Pst1= - R - 0.5 *ρ* v1

2. {8} Seefigure 4.



Accordingto {4} The value of FSP is then: FSP= Pst2- Ptot1=Pst2- (Pst1+Pdyn1),

=0 - (- R - 0.5 *ρ* v12+ 0.5 *ρ* v1

2) = R Pa. {9} Sothe absolute difference between FSP and Pst1is: ¦FSP¦-¦Pst1¦= -0.5 *ρ* v1

2 {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. Pst1 versus ambient pressure is negative and Pdyn1 positive. Theoretically the value of Pdyn1 is equal to the reduction of Pst1 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 Pst2 since Ptot1 = 0. This follows from the definition: FSP = Pst2 - Ptot1 = Pst2 = R - 0.5 * ρ * v2

2 - 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



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 Pst2(FD) = R - 0.5 * ρ * v2

2 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 (v2>v1). Consequently Pst2(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 increasing 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 FD ID 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) + ±



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(vo

2-vi2)

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

FSPc = FSP + ∆Pdiff {14}



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 airflow 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 {15} where : cl = tip clearance [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 {16} 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 related 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 {17} FSPc = Corrected FSP ∆Pi = Additional pressure drop by i. 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 = ki ∗0.5 ∗ ρ ∗ vi

2 {18} ki = 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: vf = 4Q/(π(Df-

2 – Dh2) {19}



Dh = Fan hub diameter.

Figure 3: Influences of various inlet shapes



Figure 4: Flow resistance coefficient Kobi for obstacles at inlet



Figure 5: Flow resistance coefficient Kobo for obstacles at inlet



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,



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.

∆Pwi = 0.5 * ρ * vwi2 {20}

Fan scaling rules



Scaling of the fan operation points is done according to the so called fan laws:

P :: ρ * Ut2 {21}

Q :: Ut * Df

2 {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 Cp and Cf. Cp and Cf have the following definition: Cp = FSP/(0.5*ρ*Ut

2) {23} Cf = Q/(0.25*π*Df

2*Ut) {24} Ut = fan tip speed [m/s] Ut = 0,5*ω*Df {25} ω = angular rotation speed [rad/s] ω = 2*π*RPM/60 {26} By considering Cp 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 facility 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: σ = z*c/(π*Df) {27}

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 Cp/Cf. This is the basic principle of the fan scaling rules and selection programs.



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



DP1

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, Ptot1,2 times the surface of the fan ring section.

Fax = Ptot1,2 * ¼π * Dr2 {31}

where Dr = Diameter of the fan ring

Power and efficiency



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 Tsh and angular speed ωsh. Power input:

Nsh = Tsh*ωsh

where:

Nsh= Fan drive shaftpower

[kW]

[kW]

{28}

Tsh = fan drive shaft torque [Nm]

ωsh = angular rotation speed

Effective Poweroutput

Nair= FSP*Q

[Nm]

[kW]

{29} where: As for every power transforming engine also for a cooling fan the ratio between effective power output 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 Ptot1,2 is considered. Consequently there also exist an expression for the total efficiency: � ηtot = Ptot2,1*Q/Nsh. {30} 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



efficiencies which need to be corrected as follows:

The following correction must be made for calculation of the fan shaft power. Nshaft1 = 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.

shaft = Nshaft1

* (1-KR) The following correction must be made for calculationof the fan shaft power.



Fig 1. Influence of tip clearance on fan efficiency



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 the 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 flow and a low pressure drop can only be made when the cooling fan has a free air access and a free air outlet. This makes the air cooling installation 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 sense to do it when international standards are carefully applied. Useful and practical is the Recommended Practice For Airflow 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 2 3 Class acc to DIN24166

Air Flow +/-5% +/-10% Pressure drop +/-5% +/-10% Drive power +/-8% +/-16% Efficiency -5% A-Sound Power Level

+4 dB(A)

+ 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 at 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 a microphone the (sound) pressure variations can be observed.



Like for alternating electrical current, for both the air particle velocity v (this is not the sound velocity) and the pressure variations p there has been also defined an effective value, it is the so called r.m.s value (root mean square). When the air particle velocity v and the sound pressure p is considered, always this 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.

P ~ p∗v∗F [W] {1} Because v~p, it can be stated that: P ~ p2∗F [W] {2} The range of audible pressures is very big: from 2.10-5 Pa for just audible to 200 Pa for the pain threshold. However 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 mentioneddecibels. Doing this the following quantities are derived: The Sound Power Level PWL or LW: PWL = 10 lg (P/P0) [dB] with reference value P0 = 10-12 W and The Sound Pressure Level SPL or Lp: SPL = 10 lg (p2/p0

2) [dB] With reference value p0 2.10 –5 Pa. It is understood to be correct by knowing that the reference value for the 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] {3}

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 is 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 equal like e sphere around the source or for big installations at 1 meter distance. This method is well defined 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] 63 125 250 500 1k 2k 4k 8k A-corr. [dB] -26.2 -16.1 -8.6 -3.2 0 1.2 1 -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 values as follows:

PWL(A)=10 lg (10 0.1lg Lw(A)31.5 +100.1lgLw(A)63+........ 100.1lgLw(A)8k) [dB(A)]



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 accurately predict the noise performance of a cooling fan. You must also know about the influence of the operating 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. For a 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 disturbance 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 quantity 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 Utip 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 level.The correction terms ∆dB are related to characteristic noise mechanism in an air cooling installation: The influence 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.



Typical C-values for Howden Cooling fans are: ENF 37 dB[A] ZNF 35.5 dB[A] KNF 37 dB[A] ELF 35 dB[A] ZLF 35 dB[A] KLF 35 dB[A] ELFA 33 dB[A] ZVF 33 dB[A] SX 27 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 octave 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 respect 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



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 meter 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



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 the 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<5 Do) Formulas: h = height of fan edge of fan inlet device hmin = 1.5 Di For area 1 SPL = PWL- 10 log(2πR(R+h)) For A, B and area 2 SPL =PWL –2 –10 logF + Cspl1 +Cspl2 F= 2πR2 Cspl = 2 –6.8(1- √cosα) (0°≤α≤90°) If R ≤ Do, Cspl2 = 4.(1-R/Do), else Cspl2 =0 For zone C, 1 m straight below the fan: This is a zone which is not easy to predict since there are many acoustic influence factors:

• the fan • drive equipment • reflections from the earth

As a rule of thumb can be stated that the SPL level in plane C is SPL in the position B +6 dB(A).



Total noise

General

Looking to an ACHE with a view to noise we recognise several sources of noise. Theoretically the total noise the ACHE should be calculated by adding up all individual noise sources, however the problem is that not all noise sources of the ACHE can be predicted accurately.

Examples of noise sources in an ACHE are: 1. Aerodynamic noise of the fan impeller. 2. Swirls in the air due to a non optimum fan inlet configuration. 3. Noise due to obstacles (e.g. driver support) in the air stream. 4. Noise of the E motor. 5. Noise of the power transmission (e.g. V belts, gear boxes). 6. Noise due to air passing the fin tubes. 7. Reflection (Non free field condition)

As an extend to the above there can be all kinds of elusive noise sources like noise due to the interaction between all components of the ACHE, Vibrations can be generated in the construction and can cause radiation of noise. Noise can also be amplified by the ACHE (e.g. Plenum chamber) or fan blades which are usually hollow.

The different noise sources

Impeller, inlet obstacles From the above mentioned noise sources the first three are relatively easy to predict although some impeller manufacturers will only guarantee the aerodynamic noise of their impeller in a configuration similar to their test arrangement. The noise of the support structure or inlet type noise is not always taken into account. The aerodynamic noise is calculated by Howden according a formula which is empirically defined in the course of the more than 40 years experience we have. The result of this is shown on our fan selection sheet as impeller sound PoWer Level. Beside this, a fan supplier should compare the type of inlet which he uses during the laboratory testing of his impellers with the actual inlet configuration and must consider the differences carefully. A less optimum inlet configuration in practice than during lab testing means more swirl at the fan tip area which results in a higher



power consumption and higher impeller Sound Power Level up to +3 dB(A) The support structure of the driving arrangement is positioned in the air stream. When it Is close to the impeller it's influence is bigger than when it is further away. Howden always specifies the type of inlet and obstacle data on which the given fan performance is based. The influence of the obstacles and inlet type are also specified separately.

E-Motors Besides the impeller, also E-motors and transmissions generate noise. In principle the noise levels of that equipment must be guaranteed by the suppliers. Theoretically the total noise level of the installation, i.e. the impeller and the drive is found by the logarithmic addition of the individual figures. However in practise this is not that simple because of the following reasons: · Due to the acoustic interference of components the total noise can be significant higher than the calculated value. This is the so called construction noise effect which is caused by the transmission of vibration from one component to the other. Together with the noise source the component which is affected by the vibration can generate the same frequency even at a higher level. · In most of the times the drive components are the vibration generators which transmit their vibration to the impeller or to the support structure. This is the reason that the suppliers of drive components are hardly willing to guarantee the noise level for an in situ installation, but only for on-loaded, isolated and ideal test facilities.

Power transmission Should be considered carefully in case of gearboxes or power belts. V-belts only in combination with super low noise fans.

Fin tubes, reflection and other elusive noise sources Beside the fact that air passing the fin tubes causes noise, the fin tubes can also reflect the impellers noise. Normally the Howden Impeller Sound Power level can be considered as a total noise source. When it is divided into two directions (e.g. suction side and discharge side) the PWL for one side is lower. Theoretically this should be 3 dB but how much this is in practice can also depend on the fin tube reflections. Beside the reflection of the fin tubes, ground reflections or reflection of other constructions build near the ACHE can be of influence. Specially when the guaranteed noise value is a sound pressure level at a certain position this must be considered carefully. Knowledge of local circumstances is necessary to estimate this.



What remains are all possible types of noise sources which are very elusive and difficult to predict like the generation and amplification of vibrations. Vibrations can be generated everywhere in the ACHE but they also can pass on to another part of the ACHE (e.g. plenum chamber or fan blades) and radiate noise at a completely different location than where they are generated.

Whether noise sources likes this emerges in the total ACHE noise depends on a lot of thinks like: - Distance between fan operating frequency and ACHE natural frequency (tends the ACHE to go into vibrations?). - Stiffness of the ACHE (support, plenum chamber). - Conditions of the bearings. - Alignment of the equipment. - Value of other noise source. (When the impellers and motors are very dominant you will have less nuisance of other noise sources. However their influence will be more and more in case the previous dominant noise sources are reduced). Managing these kind of noises is basically a matter of avoiding them instead of anything else. One precaution you can take is to consider the blade material you are going to apply carefully. As mentioned earlier, fan blades (both aluminium and GRP) are usually hollow and therefor very suitable to operate as a sound box for vibration generated elsewhere in the ACHE. The damping properties of the Howden glasfibre reinforced polyester (GRP) blades, is an advantage over standard aluminium blades. Beside the fact that it will give you less risk for resonance it also has less tendency to amplify vibrations generated somewhere else in the ACHE but which radiates out of the impellers blades. The Howden aluminium bladed K-series, offers a solution for both noise and vibrations due to the resilient connection of the fan blades to the hub and an elastomer buffer.

The estimation approach

ACHE data Number of bays : 5 Number of fans/bays : 2; Total number of fans = 10 Dimensions 1 bay : approx. 14.1 x 7.22 m Dimensions ACHE : approx. 14.1 x 36.1 m



Theoretically the total noise should be calculated by the energetic summation of the different noise sources of the ACHE. However not all noise sources can be predicted up front.

Our approach is to use a ACHE characteristic value to cover all the exlusive noise sources which are difficult to predict. Key question is what the value of this characteristic should be. For reasons mentioned earlier this is very dependent on the general execution of the ACHE construction, power transmission etc. but also of the total noise level of the ACHE. If only the motor and fans are dominant enough you can neglect all other noise sources.

The Howden fan selection software offers extended possibilities for the calculation of multiple fans.

Noise reduction

If expression (1), as mentioned in the "fan noise" section is accepted, it is clear that the correct approach to achieve noise reduction is to look at decreasing the characteristic value C and/or the tip speed Utip without reducing pressure drop, flow or fan efficiency. Reduction of the tip speed of a fan will indeed reduce the generated noise, however it will also reduce the pressure and flow. The reduction of pressure and flow can be avoided by making the blades wider or installing more blades, however the last way of doing enlarges the number of blade trailing edges, which means an increase of the number of noise sources. Making the blades wider is better. Wider blades perform aerodynamically the same as narrow ones, but at a lower speed. It is like a sail-plane which can fly at a lower speed than a motorized plane because of its bigger wing areas. See for instance figure 1



For the reduction of the characteristic value C, besides making the blades wider, the shape of the blade must be changed, by sweeping it forward and/or through applying the latest Aerotip technology.

Aerotip technology: The research for the Aerotip was initiated by the wish to reduce the vibration of the fan ring and supporting structure through decreasing the pressure pulses generated by the blade tips. This aim was

achieved, however the comparative pressure distribution readings indicate additional advantages; an increase of the lift generated at the blade tip resulting in a significant fan performance improvement and a noise level reduction of 1 to 2 dB(A).

Forward sweeping: This is the common conclusion of international research by several institutes over the last few decades. The most important reasons why the sweeping forward reduces the noise production of the fan are the following:

1. The forward sweep causes a phase shifting cancellation of fan noise generated at different radial stations on a blade and blade to blade interference. Here it concerns what is earlier mentioned as the noise by turbulence ingestion.

2. Forward sweep severely limits the range of spanwise travel of low momentum fluid or more simply said, it limits the grow of boundary swirls on the fan blade trailing edges by truncating the natural growing path at the trailing edge. This effect is related to the so called fan self noise.

3. The speed component of the rotating air perpendicular to the trailing edge line is significant



smaller than for a radial blade. If it is accepted that this perpendicular speed component generates the noisy swirls, its reduction must make the fan more quiet.

4. The big tip angle on the leading edge side, reduces the tip vortex shedding.

Figure 3: The typical SX shape reduces the air velocity component in perpendicular to the leading edge (vector B).

The availability of low noise and super quiet fans complicates the design of air cooling systems: Sound attenuaters are little effective, expensive as well as power consuming.

Example: For a duty point with a static pressure drop of 110 Pa and a flow of 280 m3/s, a classic 28' diameter fan performs well at 117 RPM with a PWL of 101 dB(A). A reduction of 15 dB(A) can be achieved in the PWL value by using a super low noise fan instead of a classic fan.(Fig. 4) For another application it can be interesting to reduce the number of fans without exceeding noise demands. A super low noise fan can supply 560 m3/s (twice 280 m3/s) over a resistance of 110 Pa with a noise level of 94.4 at 90 RPM. This is still 11.5 dB(A) less than two classic fans according to the first configuration. If a certain PWL value is acceptable for an application, then it is possible to operate the tower with a super low noise fan with a much bigger flow than with a classic fan. For the example of before, where the classic fan operates at 101 dB(A), a super low noise fan can supply 50 percent more air through the same cooling tower at a noise level of only 97 dB(A). This is realised with a rotation speed of 100 RPM. The conclusion is that there is a wide range of possibilities in avoiding the need for sound attenuators which demands a new evaluation approach for the cost/benefit analysis of air cooling installations.

Example: Differences in noise generation for a fan with 8 m. diameter and an air volume of 280 m3/s at a static



pressure of 110 Pa.

Figure 4: Various Cooling fan options for an identical duty point. Difference in Noise generation: up to 15 dB(A)

Figure 5: Application potential of Super Low Noise fans. Nair = Pressure * AirFlow



Mechanical integration

Proper integration of the Howden fan impeller in the air-cooled system is important to satisfactory operation of the system and its lifetime expectancy. Based on our field experience we have the following recommendations for the overall system design:

Sizing of bearings

The impeller is fitted on a drive shaft, which is part of the bearing set, gearbox or E-motor. Our recommendation is to size the support bearings for a load situation that is generated when one blade of the impeller is lost. (half a blade for a Howden super low noise SX fan).

Avoid vibration due to resonance

In order to avoid vibration due to resonance Howden fan impellers are designed to have a natural frequency, which is at least twice the value of the nominal rotation frequency of the fan.

Interference between the rotation frequency of the fan impeller and the natural frequencies of the fan drive & support structure shall be avoided at all times. Howden recommends that the natural frequencies of the fan drive & support structure are at least 1.5 times higher than the rotation frequency of the fan impeller. This needs to be checked at the design stage.

When a customer introduces a new design cooling system, or when a new fan type is applied in an existing cooling system, then interference between the rotation frequency of the fan impeller and the natural frequencies of the fan drive & support structure needs to be checked by measurements. The same for interference of multiple values of the rotation speed and natural frequencies of support structures. Special attention must be paid to interference of the blade passing frequency of the fan and the natural frequencies of the support structure. The blade passing frequency is the result of blade number of the fan and the rotation speed of the fan in Hertz. The natural frequency value of the fan impeller in operation and the blade passing frequency can be found on the output sheets of Howden's fan selection program CF-P20.

In line with our recommendation it is advised that the shaft, which is supporting the fan impeller, has at least



two bearings, which both have an independent rigid foundation, and that a single flange type of support structure is avoided. Particular attention must be paid for 3 and 4 bladed fans. Figure 1 and 2 show two examples of drive concepts: Figure 1 reviews an E-motor which is flanged on a gearbox. This concept is sensitive for various rocking modes. In figure 2, the E-motor is independent fixed to the fan bridge, which will be more stable:

Figure 1: Flanged motor, which is very sensitive for various rocking modes

Figure 2: E-motor which is independent supported on the fan bridge

Variable Frequency Drives (VFD)

A Variable Frequency Drive (VFD) is today's method for combining reduction of mechanical loading and optimal process control. A consequence of applying a VFD is that the rotation speed can continuously vary from zero to nominal value. As a consequence also the excitation frequencies like the rotation speed, the blade passing frequencies and their multiple values vary continuously.

Howden K-series aluminium fans can be used without limitation across the full range of a VFD. There is no



need to block out a range of frequencies and tip speeds due to the Anti Vibration System (AVS) blade support design. Howden polyester fans have hardly any sensitiveness for resonance. The natural frequency of the blades is always more than twice the maximum nominal speed. Therefore the only theoretical risk on resonance of the blades that remains is when the rotation speed (operation frequency= OF ) is that low that the blade passing frequency (BPF) interferes with the blade operating natural frequency. All these values can be found on the output of the CF-P20 selection program. From that value the interference point is found by taking the BONF/BPF ratio. Multiplying this value by the fan RPM gives the interfering g fan rotation speed. Multiplying the ratio BONF/BPF by the fan tip speed gives the interfering tip speed. As long as this tip speed is lower than 35 m/s, the excitation will be that low that there does not exist a risk on resonance of the blades. In other cases it is advised to consult Howden.

Example for a model 32 ENF 8 Howden Polyester fan:

BONF = 4.9 Hz OF = 2.0 Hz BPF = 15.7 Hz RPM = 117.8 Tip Speed = 60.2 m/s BONF/BPF = 0.312 * tipspeed = 18.8 m/s A more extensive analysis with Campbell diagrams is not needed for Howden Cooling Fans.

A more realistic risk on resonance and vibration problems is interference of the motor-fan unit with the support structure. Recognised excitation frequencies are the operation frequency and the blade passing frequency. Critical operation points are hardly to predict but easily to correct during commissioning. That is why it is advised to do a few test runs over the full operating range of the VFD during commissioning and to track on disturbing frequencies. With a modern VFD those frequencies can be blocked for operation.

Above recommendations are intended as a general guide when customers design the complete cooling system. As such Howden does not accept overall system responsibility or any contractual liability for vibration problems that may occur due to resonance between the fan rotation frequency and natural frequencies of the support structure as far as



components and engineering tasks are not part of Howden's supply.

Imbalance

The radial forces caused by imbalance of a cooling fan are calculated as follows: Howden cooling fan impellers are balanced to G 6.3 according to ISO 1940/1. This means that the point of gravitation of the fan has a circumferential speed of 6.3 mm/s The theoretical permanent imbalance force FG6.3 in fact is a centrifugal force. FG6.3 = m∗ω2∗e [N] Where:

m = impeller mass [kg]

find in dimension sheet

ω

= angular speed [rad/s] calculate from

RPM: ω = 2∗π∗RPM/60

e = excentricity [m] find from: e =

6.3/(ω∗1000) This force FG6.3 will have a very low value. For the dimensioning of bearings it is advised to calculate with a load situation that is generated when one blade of the impeller is lost (half a blade for a Howden super low noise SX fan). Please refer to Howden standard 03-07-341. The theoretical radial force Frad-1bl, when one blade is lost is calculated as follows: Frad-1bl = m1bl ∗ 0.125∗ Df ∗ ω2 [N] Where: Df = Fan Diameter [m] m1bl = mass of one blade [kg] m1bl can be found in the fan dimension sheet from the mass difference between two indentical impellers with one blade difference of blade number. For example the mass of one blade 12ENF is the difference between the mass of a 12 ENF 5 and a 12ENF4 impeller:



136 – 118= 18 kg

Pulse force

The pulse force on the fan casing caused by the passing of a fan blade, can be estimated as follows: Consider a fixed position on the fan casing where the blade is passing, each time when this occurs, the fan casing feels an under pressure pulse from the blade tip. The value of the under pressure is:

with ρ = the air density (kg/m3) utip = the tip speed (m/s) This pressure is effective on a certain area of the fancasing. Assuming that this area A is equal to the tip section of the blade then the radial pulse force is:

Theforce on a specific area of the fancasing is a cyclic load with a frequency of(z×n)/60 (Hz), with z = the number of blades of the fan and n = the rotationalspeed. It has a real pulse character.



The force on a specific area of the fancasing is a cyclic load with a frequency of (z × n)/60 (Hz), with z = the number of blades of the fan and n = the rotational speed. It has a real pulse character.



fan parameter

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