# Empirical rules for determining fan noise

The desire to have a simple rule by which the noise output of a fan could be deduced from its operational duty is apparent. An early attempt was made by Beranek, Kamperman and Alien, when the following relationship was proposed:

PWL =100+1 Olog HP dB re 10“13 W Equ 14.4

Where:

PWL = overall acoustic power level of noise transmit­ted along ducts fitted to the inlet and outlet of fan operating at or near its peak efficiency

 X
 +10
 0
 -10
 -30
 -40
 О СО о О in о о ю «о со со о сч ЗДч- Ч — ч — СЧ CN СО
 ООО ООО О о ю СО О
 Hz
 = nameplate horsepower of the driving motor
 -10

 -20 DB -30

 -40

 -50

 Figure 14.7 Weighted sound pressure curve A +10

 0

 -40

 -50

 Figure 14.8 Weighted sound pressure curve B 220 FANS & VENTILATION

 Equ 14.5

 Where: Q P N Then

At that time the Americans were using a different base refer­ence level and if updated for present day units, the above for­mula becomes:

PWL =91.3+10 log kW dB re 10 12 W

Which looks far less attractive and could well have been a deter­rent to its use!

It will be appreciated that this formula was of necessity approxi­mate only, and was based on a series of fans tested at pres­sures up to about 500 Pa. Subsequently, with the steady in­crease in system pressures up to 2500 Pa in many cases, a revised formula was suggested:

PWL = 100+10 log HP +10 log p dB re10~13 W Equ 14.7 where:

P = pressure (ins. w. g.)

Again in modern units this becomes:

PWL =67.3+10 log kW + 10 log pdB re10“12 W Equ 14.8

Where:

P = pressure (Pa)

Bearing in mind that there can be a considerable difference be­tween absorbed and nameplate power (especially in the case of forward curved centrifugal fans), it was also suggested that the former be inserted in the formula.

A further manipulation of the power term is possible for:

Qxp = kW 10xri%

= m3/s = Pa

= fan efficiency %

PWL = 57.3 + log Q + 20 log p -10 log rf/o dB re 10’12 W

Equ 14.9

This formula gives the total noise. Assuming that inlet and out­let noise are equal, then these would each, of course, be 3 dB less.

And there the exercise should end, for one has to say that for very large fans and for fans at pressures above 1000 Pa, the uncertainty when compared with actual noise tests can be as much ±15 dB using any of these formulae, even when the fan has been selected at its peak efficiency. This is hardly surpris­ing for whilst some fan ranges which were current in 1955 are still available, research over the past thirty years or so has meant that we now have a very much better idea of the noise generating mechanisms within fans.

Research into the cut-off and volute design of centrifugal units has, in itself, led to improvements of over 10 dB whilst in axial fans, the importance of tip clearance, impeller-casing concen­tricity, rotor-statorgap, and rotor-stator vane numbers, have all been the subject of important work.

It might be said that use of empirical formulae, such as those above, has by experience given results similar to manufactur­ers’ claims. This does not necessarily confirm their correctness

— indeed it may simply show that that particular manufacturer does not have noise measuring facilities, and therefore, uses the self-same formulae.

Noise measuring equipment and laboratories are extremely ex­pensive. It is a matter of regret that only a few of the major man­ufacturers have invested in such facilities and that many of the others continue to use such empirical formulae. The alternative is to sub-contract such sound testing to one of the many inde­pendent laboratories now capable of this.

Posted in Fans Ventilation A Practical Guide