RESEARCH ARTICLE

Jianfeng MA, Datong QI, Yijun MAO

Noise reduction for centrifugal fan with non-isometric forward-swept blade impeller

E Higher Education Press and Springer-Verlag 2008

Abstract To reduce the noise of the T9-19No.4A centri-

fugal fan, whose impeller has equidistant forward-swept

blades, two new impellers with different blade spacing

were designed and an experimental study was conducted.

Both the fan’s aerodynamic performance and noise were

measured when the two redesigned impellers were com-

pared with the original ones. The test results are discussed

in detail and the effect of the noise reduction method for a

centrifugal fan using impellers with non-isometric for-

ward-swept blades was analyzed, which can serve as a

reference for researches on reduction of fan noise.

Keywords centrifugal fan, non-isometric blade impeller,

noise reduction

1 Introduction

One of the effective methods to reduce the noise of a fan is

to control the noise sources. There are several ways to

control the aerodynamic noise sources of centrifugal fans

[1], but applying the same noise-reduction method to dif-

ferent kinds of fans may lead to different results. One

method may obtain noticeable noise reduction, while

another may show no effect at all. Therefore, it is neces-

sary that further investigations be conducted in this field.

A fan’s aerodynamic noise is often dominated by tones

at the blade passing frequency (BPF) and its higher har-

monics caused by the strong flow interaction between the

rotating impeller and stationary volute casing. To reduce

the BPF sound pressure level (SPL), the method of using

non-isometric blades has been proposed as this can dis-

perse the SPL at BPF and reduce the total SPL.

In 1967, Lowson proposed that by distributing blade

spacing unequally, the BPF sound level of axial flow com-

pressors can be reduced [2]. In 1970, Melin and Sovran also

suggested that the method of using non-isometric blades

can reduce the discrete noise level of the impeller. They

deduced an equation to compute the discrete sound level

of non-isometric blades [3]. Later, Ewald [4] adopted some

similar noise-reducing methods by which the BPF sound

level of a 22-blade axial flow fan was decreased by about

8 dB. Other researches conducted by Duncan also corrobo-

rated this method of reducing the sound level by spacing

blades unequally on the impeller of an axial flow fan [5]. In

1980, using the blade spacing model, Krishnappa1) theoret-

ically deduced governing equations of sound radiation, con-

sidering non-isometric blades for the whole fan, and

conducted several experiments. In China, Sun Xiaofeng

discussed the influence of a non-isometric blade on the

BPF sound level at the observation point of an axial flow

fan [6]. Due to developments in aeronautics, earlier

researches mostly focused on the axial flow fan using a

non-isometric blade, but few researches were conducted

on a centrifugal fan with forward-swept blades.

In this paper, based on previous studies which intro-

duced non-isometric blades for centrifugal fans, the effects

of two types of different blade spacing impellers on aero-

dynamic performance and aerodynamic noise in a centri-

fugal fan with forward-swept blades were compared and

analyzed. This centrifugal fan used had 12 forward curved

blades driven by a 3.0 kW AC motor, rotating at 2900

r/min. It consisted of an impeller, a vaneless diffuser,

and a volute. The blade angles at the inlet and outlet were

38u and 126u, respectively, and the inlet and outlet dia-

meters of the impeller were 156 and 400 mm, respectively.

The vaneless diffuser, whose outlet diameter is 460 mm,

rotated with the impeller.

Translated from Fluid Machinery, 2007, 35(9): 1–4, 37 [译自: 流体机

械]

Jianfeng MASchool of Energy and Power Engineering, Xi’an JiaotongUniversity, Xi’an 710049, China;Shenyang Blower Works Group Co., Ltd, Shenyang 110142, ChinaE-mail: jfma218@yahoo.com.cn

Datong QI, Yijun MAOSchool of Energy and Power Engineering, Xi’an JiaotongUniversity, Xi’an 710049, China

1)Krishnappa G. Effect of modulated blade spacing on centrifugal fan noise. Proceedings of Inter-Noise’80, 1980, 215–218

Front. Energy Power Eng. China 2008, 2(4): 433–437DOI 10.1007/s11708-008-0082-6

2 Design method of non-isometric bladeimpeller

Two constraints should be considered when designing a

non-isometric blade impeller. One is that after modifying

the blade spacing, the dynamic balance condition should beguaranteed when the impeller is running. The other is that

the variation of the angle between blades should not be too

large; otherwise it may cause a large deterioration of aero-

dynamic performance. The circumferential angles of blades

with respect to the original and redesigned fans are denoted

by h and h9. Assuming that the first circumferential angle of

the blades is h1~h01~00, the other circumferential angle of

the blades are [h2,h3,…,hN] and h02,h

03, . . . ,h

0N

h i. Since the

tested fan is running at low speed, the influence of aero-

dynamic load distribution on the dynamic balance of the

impeller is ignored. Thus, the mathematical representation

of the dynamic balance constraint is

1zPNi~2

cos h0i~0,

PNi~2

sin h0i~0:

9>>>=>>>;

ð1Þ

The geometric constraint condition on the circumfer-

ential angle interval between two adjacent blades is

2p

N{af h

0iz1{h

0i

������f 2p

Nza, ð2Þ

where N is the blade number. The selection of angle adepends on different conditions, and in this paper, a5 5u.

Many different schemes of combining blades can be

derived from meeting the constraint conditions, i.e.,Eqs. (1) and (2). In this paper, the blade grouping scheme

detailed in Ref. [6] is adopted. In this scheme, the circum-

ferential angle of blade No. i of the redesigned impeller is

defined as h0i~hizA sin shið Þ, where s is the number of

groups, A is the phase-modulated quantity. The com-

pound mode of a blade on the T9-19No.4A centrifugal

fan was redesigned according to the above formula. To

keep the aerodynamic performance of the fan almostunchanged, the circumferential angle interval of the

blades was designed to be between 25u and 35u. The angleinterval between two adjacent blades is Dh

Dh~h0iz1{h

0i

~ hiz1{hið ÞzA sin shiz1ð Þ{ sin shið Þ½ �

~300z2A sins|300

2cos

s 2hiz300ð Þ2

: ð3Þ

The above formula should satisfy

A sins|300

2cos

s 2hiz300ð Þ2

��������v2:50: ð4Þ

Hence, the span of phase-modulated quantity A is

obtained by using the above constraint condition, as

shown in Table 1.

3 Selection of design scheme

According to the design method described in the previous

section, two different types of blade design scheme are

employed. For convenience, the unmodified fan is called

the original fan and the two redesigned fans are called

redesigned-fan I and redesigned-fan II, respectively.

Table 2 shows the specific parameters of these two design

schemes, wherein a is the circumferential angle, and Da is

the circumferential angle interval. Due to the limitation of

machining accuracy and processing conditions in the fan

factory, it is inevitable that there exists some deviation in

the impeller dimension from that in the drawing paper.

Both the design values and the actual dimensions of the

redesigned impellers are shown in Table 2.

4 Experimental installation and results

A test installation for the aerodynamic and acoustic char-

acteristics of the investigated fan was designed according

to China standards GB/T2888-91: Methods of Noise

Measurement for Fans, Blowers Compressors and

Roots Blowers, and GB/T1236-2000: Industrial Fans-

Performance Testing Using Standardized Airways [7,8].

Figure 1 is the illustration of this test installation with

its main elements. In order to compare the original impel-

ler with the redesigned ones, the same volute, air inlet,

inlet duct, electric motor and measurement system are

used. Hence, the only difference between the original

and redesigned fans is the use of different impellers.

The performance curves and the SPL spectra are

obtained in a half anechoic chamber. In Figs. 2 and 3,

the A-weighted SPL and specific A-weighted SPL are

represented. The total pressure and the efficiency are

represented as a function of the volume flow, as shown

in Figs. 4 and 5. The SPL spectra for the investigated fans

operating at three volume flow rates are shown in Figs. 6–

8. The test results show that the aerodynamic perfor-

mances of the two redesigned fans are similar, but the

noise of redesigned-fan II is weaker than that of rede-

signed-fan I. Thus, only the comparison between the rede-

signed-fan II and the original fan is presented.

The redesigned-fan II attains its maximum efficiency

when the volume flow is about 17 m3/min. In the

Table 1 Span of phase-modulated quantity

s 1 2 3 4

span of A/(u) 0–9.6 0–5 0–3.5 0–2.8

434 Jianfeng MA, et al.

A-weighted SPL spectra, there is a conspicuous peak at

BPF (i.e., at 595.66 Hz), which is slightly lower than that

of the original fan. At 749.89 Hz, a new peak emerges,

while there is no such evident one in the SPL spectra of

the original fan. The deviation of the total SPL between

redesigned-fan II and the original fan is small. In small

volume flow (about 7 m3/min), the SPLs of the original

fan at BPF and 668.34 Hz are almost the same. The SPLs

Table 2 Parameters of two types of design scheme

No. original fan redesigned-fan I (s5 2,A5 5) redesigned-fan II (s5 3,A5 3.5)

a/(u) design value actual value design value actual value

a/(u) Da/(u) a/(u) Da/(u) a/(u) Da/(u) a/(u) Da/(u)

1 30 34.3 34.3 35 35 33.5 33.5 32.5 32.5

2 60 64.3 30 63.5 28.5 60 26.5 58.5 26

3 90 90 25.7 88 24.5 86.5 26.5 84.5 26

4 120 115.7 25.7 116 28 120 33.5 118 33.5

5 150 145.7 30 144.5 28.5 153.5 33.5 151.5 33.5

6 180 180 34.3 178 33.5 180 26.5 179 27.5

7 210 214.3 34.3 213.5 35.5 206.5 26.5 206 27

8 240 244.3 30 243.5 30 240 33.5 238.5 32.5

9 270 270 25.7 270.5 27 273.5 33.5 272 33.5

10 300 295.7 25.7 295 24.5 300 26.5 298.5 27

11 330 325.7 30 324.5 29.5 326.5 26.5 325 27

12 360 360 34.3 358.8 35.5 360 33.5 359 34

Fig. 2 A-weighted SPL vs volume flow Fig. 3 Specific A-weighted SPL vs volume flow

Fig. 1 Test installation

Noise reduction for centrifugal fan 435

of the redesigned-fan II at BPF and 749.89 Hz are

obviously lower than that of the original fan. Compared

with the original fan, the total A-weighted SPL of rede-

signed-fan II is reduced by 1 to 2.5 dB at this operating

mode. At an even larger flow condition (34 m3/min), in the

entire SPL spectra of the original fan, the main contribu-

tions are noise whose frequencies are BPF at 578.76 Hz,

whose second harmonic is at 1154.78 Hz and whose third

harmonic is at 1727.83 Hz, respectively, of which the BPF

SPL is the maximum. Compared with the original fan, the

total A-weighted SPL of redesigned-fan II increases by 1

to 4 dB at this operating mode. In addition, as far as

aerodynamic performance is concerned, the total pressure

of redesigned-fan II drops a little over the entire operating

mode, as does its efficiency.

5 Conclusions

The method of noise reduction for a centrifugal fan with a

non-isometric forward-swept blade impeller is studied.

Some key conclusions of this investigation are summar-

ized as follows.

1) In this paper, the attempt to adopt a non-isometric

blade impeller for a centrifugal fan is not successful.

Fig. 6 Comparison of measured A-weighted SPL at threevolume flow (original fan)

Fig. 4 Total pressure vs volume flow

Fig. 5 Efficiency vs volume flow

Fig. 7 Comparison of measured A-weighted SPL at threevolume flow (redesigned-fan II)

Fig. 8 Comparison of measured A-weighted SPL at threevolume flow (redesigned-fan I)

436 Jianfeng MA, et al.

Compared with the original fan, at the domain of larger

flow, the sound level of the redesigned fan is increased by

about 1 to 4 dB, with the maximum efficiency operating

mode and the sound level of the redesigned fan almost

remaining unchanged or slightly reduced. However, atsmall flow, the sound level of the redesigned fan can be

reduced by about 1 to 2.5 dB, which indicates that the

possibility of fan noise reduction does exist. Further

investigations and studies should be conducted on the

adoption of a non-isometric blade impeller.

2) The test indicates that by using a non-isometric blade

impeller, the sound levels of certain frequencies of the

centrifugal fan with a forward-swept blade can be chan-ged, which provides the possibility of noise reduction and

sound quality improvement for a centrifugal fan with a

forward-swept blade.

3) There exists some deviation in impeller dimension

from that in the drawing paper; hence, it is difficult to

analyze the regularity of some experimental phenomena.

More attention should be paid to working on accuracy in

future research.4) There are only two improvement schemes in this test.

More improvement schemes should be attempted to

explore more appropriate distribution regularities of a

non-isometric blade to obtain more believable analyses

and conclusions.

5) The test indicates that changing the isometric blade

into a non-isometric blade may lead to deterioration in the

aerodynamic performance of the fan. It is suggested thatthe blade profile be modified, with changes in the circum-

ferential spacing of blades, so as to explore the best design

method for blade profiles while using a non-isometric

blade. This can both keep better flow in the blade channel

of the impeller and maintain the aerodynamic perform-

ance of the fan. Moreover, it may be of benefit for dis-

playing the noise reduction effect of non-isometric blades.

6) The fan investigated in this test is a T9-19No.4A

centrifugal fan. The method described in this paper is

mainly for small types of forward-swept centrifugal fans

which have a higher pressure coefficient and a smaller

flow coefficient. It is hoped that this paper can serve asa reference for research in similar products.

References

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3. Mellin R C, Sovran G. Controlling the tonal characteristics ofaerodynamic noise generated by fan rotors. Journal of BasicEngineering (Series D), 1978, 72: 143–154

4. Ewald D, Pavlovic A, Bollinger J G. Noise reduction by apply-ing modulation principles. Journal of the Acoustical Society ofAmerica, 1971, 49: 1385

5. Duncan P E, Dawson B. Reduction of interaction tones fromaxial flow fans by suitable design of rotor configuration.Journal of Sound and Vibration, 1974, 33(2): 143–154

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