Chen Wang1

Laboratory of Aerodynamics and Acoustics,

HKU Zhejiang Institute of

Research and Innovation and

Department of Mechanical Engineering,

The University of Hong Kong,

Pokfulam, Hong Kong

e-mail: chadwong@hku.hk

Lixi HuangLaboratory of Aerodynamics and Acoustics,

HKU Zhejiang Institute of

Research and Innovation and

Department of Mechanical Engineering,

The University of Hong Kong,

Pokfulam, Hong Kong

e-mail: lixi@hku.hk

Passive Noise Reductionfor a Contrarotating FanThere has been renewed interest in the contrarotating (CR) fan configuration in aviationand other applications where size and weight are important design factors. Contra-rotation recovers swirl energy compared with the single-rotor design, but this advantageis not fully harnessed due to, perhaps, the issue of noise. This study explores passive noisereduction for a small, axial-flow, CR fan with perforated trailing-edge for the upstreamrotor and perforated leading-edge for the downstream rotor. The fan is designed withsimple velocity triangle analyses, which are checked by 3D flow computations. The aero-dynamic consequence and the acoustic benefit of such perforated blading are investigatedexperimentally. The results show that there is a reduction of total pressure comparedwith the baseline CR fan at the same rotating speeds, but this is easily compensated forby slightly raising the rotating speeds. A reduction of 6–7 dB in overall noise is achievedfor the same aerodynamic output, although there is a moderate noise increase in the highfrequency range of 12.5–15.0 kHz due to blade perforations. The effect of inter-rotor sep-aration distance is also investigated for the baseline design. A clear critical distanceexists below which the increased spacing shows clear acoustic benefits.[DOI: 10.1115/1.4028357]

1 Introduction

The aviation industry has been continuously trying to increasethe efficiency and reduce the environmental impact of aero-engines in order to cope with the increasing fuel prices and risk ofoil shortage and meet the requirements of ever more stringentlegislation on emitted pollutants and noise. The fuel consumptionof jet engines has decreased about 40% since 1960s. This reduc-tion has been accomplished by the continuous improvement ofengine technology, leading to increased component efficienciesand the change from turbojet to turbofan engines with increasingbypass ratios [1].

1.1 CR Turbomachinery. With respect to the future engineconfiguration, there has been renewed interest in the CR turboma-chinery components including propellers, fans, compressors, andturbines since such configuration has potential advantages overthe conventional design. A CR configuration can decrease the sizeand weight of an aero-engine by eliminating stators, thus leadingto an increased efficiency and thrust-to-weight ratio. In addition,the gyroscopic moment of the aero-engine can be decreased byutilizing CR shafts. Furthermore, the ability to achieve high pres-sure ratios with decreased rotating speeds can reduce engine noiseand this can be significant in the case of CR fans [2]. The mainaerodynamic advantage of CR configuration stems from recoveryof the swirl velocity losses of the front rotor by the aft rotor. Thefront rotor imparts a tangential velocity to the air as it passes by.This swirl velocity acts as an additional angular velocity for theaft rotor, without the power plant having to drive the aft rotor at ahigher angular velocity [3].

The review of Mitchell and Mikkelson [4] suggested that pro-pulsive efficiency could be increased by 7–11% by introducingCR propeller compared with an equivalent single-rotating propel-ler. Strack et al. [5] substantiated this suggestion through an ana-lytical study which took into account performance, acoustics,vibration, weight, cost, and maintenance. Study by Bradley [6]showed that turboprops with these CR propellers can give specific

fuel consumption improvements of 25–30% over an equivalenttechnology turbofan.

Young [7] proposed the CR configuration for compression sys-tems in early 1950s through studying a fan stage comprising ofCR rotors, which showed higher pressure rise and swallowingcapacity. The potential benefits of CR fans were also recognizedand explored in the work of other scholars [8–11]. Shigemitsuet al. [12–14] studied the performance and the internal flow condi-tion of a CR small-sized axial fan at the designed flow rate andthe partial flow rate. They also clarified the unsteady flow condi-tions at the inlet and the outlet of each rotor with unsteady numer-ical results [15]. An experimental study on CR axial-flow fanswas carried out by Nouri et al. [16,17]. The results showed thatthe efficiency was strongly increased compared to a conventionalrotor or to a rotor–stator stage. The effects of varying the rotatingspeed ratio and inter-rotor axial distance on the overall performan-ces were also studied. The performance and detailed flow struc-ture of a CR compressor under different rotating speeds andtypical working conditions were experimentally and numericallyinvestigated by Chen et al. [18]. For fans and compressors, thehigh diffusion and issues with turbulence encountered in CRdesigns is one of the reasons why most recent works employ flowcontrol mechanisms, such as boundary layer aspiration and split-ters in order to have acceptable losses [2].

CR turbines represent the state of the art of actual and futureaero-engines, built for a considerable reduction of weight and fuelconsumption. In the 1950s, Wintucky and Stewart [19] predictedan overall efficiency increase due to the elimination of the inter-stage stators through studying a two-stage CR turbine. Louis [20]compared CR turbines with a conventional high reaction turbineand revealed that similar stage loading coefficients can beachieved with higher efficiency when utilizing the CR concept.Lengani et al. [21] conducted an experimental investigation on theunsteady interactions in a two-spool CR transonic turbine, whichwas of limited number in the open literature. The test setup con-sisted of a high pressure stage, a diffusing midturbine frame withturning struts and a shrouded low pressure rotor. Zhou et al. [22]performed the aerodynamics design of a two-stage vaneless CRturbine and discussed the optimal selection of velocity trianglesbased on theoretical analysis. Besides, the CR concept was alsoapplied in wind turbines in order to extract more energy than thetheoretical Betz limit from wind. Appa and Forest [23] suggested

1Corresponding author.Contributed by the International Gas Turbine Institute (IGTI) of ASME for

publication in the JOURNAL OF TURBOMACHINERY. Manuscript received July 30, 2014;final manuscript received August 6, 2014; published online September 30, 2014.Editor: Ronald Bunker.

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that using a CR rotor system could increase power conversion effi-ciency of wind turbine by 40%.

1.2 General Rotor Noise Mechanism. A general expressionfor the sound field of a point force in arbitrary motion was foundby Lowson [24] in 1965, followed by the expressions for thesound fields of a point acoustic stress in arbitrary motion. Theresearch on small fan aero-acoustics was reviewed by Huang [25].The general field of fan aero-acoustics spans from Gutin’s effortin quantifying the steady loading noise in 1936, later known as theGutin noise, Lighthill’s acoustic analogy for aerodynamic soundin 1952, to Ffowcs Williams and Hawkings equation (1969)accounting for the effect of solid boundaries in arbitrary motion.

According to Blake [26], sound from rotors may be usefullyclassified as self-noise and interaction noise. Interaction noisemeans all sounds that result from the encounter of a rotating bladewith a time-varying disturbance in a reference frame moving withthe blade element. Self-noise means sound resulting from flowover the blades themselves requiring no steady or unsteady inflowdistortions. Gutin noise is a form of self-noise, which is propor-tional to the steady loading on the rotor. Important causes of inter-action noise are: rotor–stator interaction in turbomachinery; bladetip vortex interaction in helicopter acoustics; inlet flow disturban-ces caused by vortical flows and large-scale turbulence in forwardstators, rotors, and grilles; the interaction of rotor blades withannular boundary layers in ducted rotors. Blake deduced the gen-eral form of acoustic spectrum formula of rotor blade forcesresulting from inhomogeneous inflow from theoretical free-fieldacoustics of rotors. For example, concerning the sound radiationdue to the interaction of moving blades with the upstream blades,their interaction can be regarded as the incident gust of velocity(velocity defect or wake) caused by the viscous and potentialwakes of the upstream blades impinging on the downstreamblades. In general, this incident flow consists of a steady compo-nent on which both periodic (deterministic) and turbulent compo-nents are superimposed. A theory was also presented for thediscrete-frequency sound radiated by axial-flow fans and compres-sors in the work of Lowson [27] in 1970. The theory was based onthe noise radiation from the fluctuating forces on either a rotor ora stator stage due to interactions with upstream components. Oncethe wake geometry was defined, the theory enabled one to performcalculations of the noise observed at any point and the resultagreed well with experiment. An analytically based model wasdeveloped by Cooper and Peake [28] to study and predict rotor–stator interaction noise in aero-engines, in particular upstream-radiated noise, which included the important effect of meanswirling flow on both the rotor wake evolution and the acousticresponse.

The prediction of noise of traditional turbomachinery configu-ration characterized by the rotor–stator stage has reached a rela-tively high standard as a result of much theoretical work and noisetests of models. In contrast, the theoretical research into the noiseof CR turbomachinery can only date back to the 1980s. However,in spite of the late start, there is a significant amount of literatureon that subject until now perhaps due to the revived interest in CRconfiguration led by the increasingly stringent environmental reg-ulation. A theoretical prediction scheme was developed for thetone noise generated by a CR propeller with asymptotic approxi-mation techniques in Parry’s Ph.D. thesis [29] in 1988. In additionto the theoretical formula, other literatures include experimentalmeasurement, numerical computation, and empirical predictionresults to address the noise of CR turbomachinery. Bradley [6]performed a series of noise measurements on a CR propellerdriven aircraft. Both near and far-field data were obtained fromwhich it is possible to extract the tonal noise, resulting from theaerodynamic interaction between the two propeller rows by a tonesplitting technique. Shin et al. [30] investigated the rotor wake/vortex flow field generated in a CR unducted fan engine usingthree-dimensional hot-wire anemometry. Their work provided a

set of benchmark experimental aerodynamic data defining therotor wake and vortex structure, particularly in the tip region,which can relate this observed flow structure to its acoustic signa-ture. Polacsek and Barrier [31] performed an aero-acoustic simu-lation of interaction noise generated by a CR fan model, whosemain objective was to check the ability of advanced computationsto give a reliable evaluation of noise generation, propagation andradiation. An original and fast semi-empirical method was pro-posed by Lewy [32] to predict radiated sound levels for CR openrotors. Envia [33,34] summarized the recent results from theNASA’s effort to validate an open rotor noise prediction codewhich is based on a high-blade-count asymptotic approximationto the Ffowcs–Williams–Hawkings equation with the unsteadyaerodynamic simulation results as the input. The results suggestedthat the noise trends were reasonably well predicted by thisapproach. Peake and Parry [35] reviewed current scientific andtechnological issues in the quest to reduce aero-engine noise,including modern turbofans and fuel-efficient open rotors andthey also described a number of novel design modifications inopen rotor engine design for low noise.

1.3 Interaction Noise Abatement. Progress in interactionnoise reduction of blade rows has been revolutionary in the pastfew decades through the classical methods including an appropri-ate choice of rotor–stator blade count and increase of the rotor–stator axial distance, which allows the wakes generated by therotors to decay and dissipate before impinging on the stators. Cri-gler and Copeland [36] conducted an experimental study of asingle-stage axial-flow compressor to investigate the effect ofinlet-guide-vane–rotor interaction on the noise radiation patternswith a view toward alleviating the noise at its source. The resultsshowed how the radiation patterns were affected by the relativenumber of rotor blades and guide vanes. An increase in axialspacing of rotor and inlet guide vanes gave large reductions innoise levels while spacing had little effect on the overall noiseradiation directivity patterns. Dittmar [37] proposed a concept fora CR fan with reduced tone noise. A CR fan with 106 blades ineach rotor was investigated in this report. Increasing the bladenumber of the fan shifted the tones to a higher frequency wherethey were not weighted as strongly in the perceived noise levelcalculation. The additional concept in this report was to use a CRfan to shift the cut-on tone at twice the blade passing frequency(BPF) to a frequency outside the rated range.

New approaches were mainly boundary layer suction (BLS)and trailing edge blowing (TEB) as two hot areas of research inrecent years, which were removal or addition of fluid to reducethe velocity defect of blade wakes. Guillot et al. [38] tested fourdifferent TEB designs including trailing edge jets, trailing edgeslots, vortex generating jets, and suction side jets on a two-dimensional rotor geometry. The results showed that the wakecould be significantly decreased with trailing edge jets and suctionside jets while the other two did not perform as well. Carter [39]incorporated both suction side jets provided by a single supplypressure source and BLS from an ejector pump on a high-turningcompressor stator. His experiments proved that using 1.6% of thetotal mass flow, the total pressure loss coefficient was reduced by65%. Naumann and Corcoran’s tests [40] showed that discrete jetblowing from the trailing edge was the most successful way toattenuate the wake on a simulated blade after examining severaldifferent configurations for TEB, including a continuous slot atthe trailing edge, a set of discrete jets, and a set of discrete jetswith vortex generators. Leitch et al. [41] assessed the effect ofTEB on four upstream stators, which achieved noise reductions of8.9 dB, 5.5 dB, and 2.6 dB, respectively in the blade passing toneat corresponding three test speeds of 30,000; 50,000; and70,000 rpm. In addition, TEB reduced the overall sound pressurelevel (SPL) in every case. The addition of TEB from the fourupstream stators did not change the operating point of the fan, andthe mass flow added by the blowing was less than 1% of the fan

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mass flow rate. But such new approaches have several problems.First, the complex inner chamber structure for air flow makes theprocessing much more difficult and requires a considerable bladethickness. Besides, they require additional auxiliary equipment,that is, an air flow jet for TEB and a pump for BLS, respectively.

It is worth mentioning a related piece of work conducted byBae and Moon [42]. They numerically investigated the noise sup-pression potential of structural perforation in the rear portion of aflat plate with blunt trailing edge and vortex shedding. Interestingresults were obtained but the acoustic mechanism treated was thescattering of boundary layer instead of the rotor–rotor (stator)interaction noise for the current study.

2 Design of the CR Fan

A small axial flow CR fan was designed with the velocity trian-gle method. The design was initiated with traditional one- andtwo-dimensional analyses. The overall design parameters and thedesign parameters of each rotor are given in Table 1. This fan wasdesigned at standard atmospheric condition to produce 5 m3/minairflow and a 50 Pa total pressure rise. The hub radii, tip radii, andaxial chords of both rotors were selected according to the size ofthe existing available motors on hand. The blade numbers foreach rotor were appropriately chosen as 7 and 5, respectively,based on Huang’s [25] elaborate discussion about the effect ofstrut number on radiated noise of small cooling fans. The selectionof other design parameters fully incorporated Talbotec and Ver-net’s work [43], which discussed a few key design parameters ofCR fans, focusing on the difference in contrast with conventionalfans, including the rotating speed ratio of front rotor to rear rotor,the axial spacing between two stages, solidity and blade count.

The velocity triangles of the midspan section were determinedby some important empirical parameters like reaction degree.Then the simplified radial equilibrium equation was solved todetermine the velocity triangles of other designate sections. Thetwo-dimensional blade profile was defined with the camber curveand two side curves, that is, suction and pressure side curves. Theside curves were designed with reference to the NACA four-digitseries airfoils. The three-dimensional blade was stacked in thespanwise direction with the center of gravity as the stacking point.Pictures of the two rotors are shown in Fig. 1.

3 Numerical Simulation and Experimental

Measurement

3.1 Numerical Simulation. A three-dimensional, steady-flow, numerical simulation of the designed CR fan was carried outwith the computational fluid dynamics software FLUENT. The com-putational domain and the mesh details in the front rotor domainare given in Fig. 2. The inlet (marked with boundary “1”) wasabout 5Ca from the leading edge of the front rotor while the dis-tance between the trailing edge of the rear rotor and the outlet(marked with boundary “2”) was about 9Ca to prevent boundary

reflection from contaminating the results. The axial separation dis-tance between the two rotors was 30 mm. In consideration of thecomplexity of the three-dimensional blades, unstructured meshcells were constructed by GAMBIT and the mesh was refinednear the regions of both rotors. The total number of the mesh ele-ments was around 3� 106. Only one passage was computed forboth rotors to save CPU time and memory. Thus, the two sides(marked with “4”) of the computational domain were set as rota-tional periodical boundary conditions. The inlet was given thetotal pressure and total temperature as design point and the outletwas set as static pressure, that is, back pressure. The interfaces(marked with “3”) between two rotors were treated as the mixingplane model.

The fluid regions of the front and rear rotors were set in two dif-ferent rotating reference frames, rotating at 3500 rpm along þZdirection and 3000 rpm along �Z direction, respectively. The huband blade surfaces were set as no slip stationary walls relative tothe corresponding rotating reference frames. The tip clearancesfor the two rotors were both 1.5 mm. The shroud surfaces were setto be a stationary wall in the absolute frame.

First, the flow was computed for 3000 steps with the gaugeback pressure of �1000 Pa to obtain a relatively reasonable flowfield. Then, this result was used as the initial flow field for otherback pressure cases. Five different gauge back pressure cases, 0,10, 20, 30, 40 Pa were performed with RNG k–e turbulence model.The residuals of all cases fell below 10�5.

Since the two rotors both stretched from 25 mm to 58.5 mm inthe spanwise direction, the relative velocity vectors on sections ofr¼ 30, 35, 40, 45, 50, and 55 mm were checked, respectively. Theresults showed that the flow field was organized well without flowseparation. Figure 3 shows the relative velocity vectors on theblade surfaces colored by relative velocity magnitude (m/s) forgauge back pressure 0 Pa.

Cases with higher gauge back pressure were also performed.However, their computational results are not shown or discussedhere for the following reasons. When the gauge back pressure waslarger than 60 Pa, the residuals could only drop to between 10�2

and 10�3. The convergence curves of mass flow rate at outletshowed dramatic fluctuations with the iteration count. Relativevelocity vectors and static pressure contours indicated that thereexisted large flow separations in both domains of the front andrear rotors. This turning point (near 50 Pa) was likely to be relatedto the design pressure rise, which was 50 Pa.

It is known that all turbulence models have their own empiricalmodel constants. For RNG k–e model used in the simulationabove, C1e¼ 1.42, C2e¼ 1.68, and Cl¼ 0.0845. In order to testthe effect of their values on the simulation results, these three con-stants were adjusted in their own specific ranges (C1e: 1.32–1.52;C2e: 1.58–1.78; and Cl: 0.0745–0.0945). Comparison of resultsindicated that this simulation case was not sensitive to the varia-tion of the model constants as long as they were kept in the rea-sonable range. In addition, the simulation results with differentturbulence models in FLUENT including Spalart–Allmaras model,

Table 1 Design parameters

Pt (Pa) Tt (K) Q (m3/min) DPt (Pa)

101,325 288.15 5 50

Front rotor Rear rotor

Rt (mm) 58.5 58.5Rh (mm) 25 25X (rpm) 3500 3000B 7 5D (mm) 120 120Work distribution 55% 45%Ca (mm) 22 22Other constraints Axial inflow Nearly axial exit-flow

Fig. 1 CAD pictures of front (left) and rear (right) rotors

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standard k–e model, RNG k–e model, realizable k–e model, stand-ard k–x model, and SST k–x model were compared as well. Theresults showed that these models all have basically the same staticpressure contours, relative velocity vectors at the section ofr¼ 40 mm, and almost the same mass flow rate and total pressurerise.

3.2 Experimental Measurement. The aerodynamic perform-ance of this designed CR fan was studied in a ducted fan test rig,which was built strictly according to ANSI/AMCA 210 standard

[44]. This standard provides rules for testing fans, under labora-tory conditions, to provide aerodynamic performance ratinginformation.

Among the methods provided in the standard, a simpler onewas selected as indicated in Fig. 4. The fan tested was installed atthe inlet of the tube whose inner diameter was the same as that ofthe fan, which was D¼ 120 mm. Otherwise, a transformationpiece was required to serve as a transition from the diameter ofthe tested fan to that of the tube. A bell-mouth was flush-mountedat the inlet of the fan which can ensure more uniform incomingflow. In addition, a flow straightener was mounted at the position

Fig. 3 Relative velocity vectors on blade surfaces

Fig. 2 (a) Computational domain and (b) mesh details in the front rotor domain

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of 3.5D downstream of the fan outlet to eliminate the nonaxialflow components and to homogenize the out-going flow. Thus, thepitot tube, which was used to measure the dynamic pressure Pv

and static pressure Ps, can obtain more reliable results down-stream. The pressure losses due to straightener and duct wall fric-tion were also taken into consideration according to someempirical expressions provided in the standard. At the other endof the tube, a throttling device was adopted to adjust the backpressure. Two thermometers were placed at the outlet of the fanand near the position of the pitot tube, respectively, to measuretheir local temperatures. Six blocks of different thickness (4, 9,14, 19, 24, and 30 mm) were made to adjust the axial spacingbetween the two rotors.

The characteristic curve (also called P–Q curve) of the fan atthe design rotating speeds, measured with the fan test rigdescribed above, is shown in Fig. 5 marked with triangles. In addi-tion, the three dimensional Reynolds-averaged numerical simula-tion results with RNG k–e model of the fan are shown in thisfigure as well, marked with circles. The five circles representcases of five different gauge back pressures, 0, 10, 20, 30, and40 Pa, from right to left side as described in Sec. 3.1. The compu-tation and experiment agreed well in the segment of 0–40 Pa.

4 Noise Reduction

4.1 Contrarotation Interaction Noise. Figure 6 illustratesthe general picture of rotor–rotor interaction for a CR fan. Accord-ing to Polacsek and Barrier [31], the interaction frequencies f12

between two CR blade rows, and the corresponding spinning cir-cumferential mode m, are addressed by the following expressions:

f12 ¼ m1B1X1 þ m2B2X2j j; m ¼ m2B2 � m1B1

where m1 and m2 are positive or negative integers, X1 and X2 aretheir rotating speeds, respectively. The sign of m is determinedwith the second row as the reference. Lewy [32] pointed out thatm1 and m2 always have the same sign because the key parameterfor sound radiation efficiency, the tip phase rotation Mach numberM/, is much larger than the tip rotational Mach number Mrot ifthey have the same sign and is much smaller than Mrot if they areof opposite sign.

In fact, the conventional expression of f12 and m for rotor–statorinteraction can be generated by setting X2¼ 0,

f12 ¼ m1B1X1j j; m ¼ m2B2 � m1B1

From the comparison, we can see that rotor–stator interactioncan be just regarded as a special case of contrarotation interaction.They have the same spinning circumferential mode. The differ-ence lies in the upstream blade row encounter with the down-stream one with a higher frequency for the contrarotationinteraction due to their counter rotation against each other. There-fore, despite late start of theoretical research into noise of CR tur-bomachinery, it can be understood as a general case of themechanism of rotor–stator interaction.

Blake [26] described the general qualities of rotor–stator inter-action noises in his book. It may be regarded as arising from cir-cumferential spatial filtering in which samples of wake harmonicsfrom the upstream components are made by the downstream blad-ing system at integer multiples of the downstream blade number.By such filtering, the downstream blading selectively responds toparticular circumferential harmonics in its flow. The circumferen-tial variations may be called velocity defects. Tinetti et al. [45]elaborated the mechanism of rotor–stator interaction noise. As arotor wake passes by and impinges on a stator, the effective angleof attack and the velocity of the relative flow change, producingtransient fluctuations in the pressure field acting on the stator vanesurfaces. This unsteady pressure field generates a lift force thatfluctuates with a frequency equal to the blade passing rate, whichgives rise to a dipole-type noise source. They also mentioned thatfan noise reduction can be achieved either by reducing the ampli-tude of the wakes shed by the upstream rotors or by reducing theresponse of the downstream stators to impinging wakes.

4.2 Blade Perforation. As discussed in Sec. 1.3, two recenthot topics for interaction noise reduction, TEB and BLS, are bothbased on reducing the velocity defect of upstream blade wakesthrough wake filling with additional high-velocity fluid or removalof the low-velocity fluid in the boundary layer. But they both havetheir own complications and limitations, which have been dis-cussed before. In the following work, we explore the effectivenessof using blade perforation to suppress the unsteady interactionforce and hence noise radiation. We use perforated trailing edgefor the upstream rotor and perforated leading edge for the down-stream rotor. Passive blade perforation can redistribute pressureon the outer surface of blades by establishing communicationbetween the pressure and suction sides of the blade. For the

Fig. 4 Fan test rig

Fig. 5 The characteristic curves of experiment andcomputation

Fig. 6 General picture of rotor–rotor interaction

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upstream rotor, trailing edge perforation may reduce the effectiveblade chord length and smear the wake velocity defects impingingupon the downstream blades. For the downstream rotor, properperforation at its leading edge region may reduce the gust loading

caused by the coming wakes. It may also be understood as havingan effective separation distance larger than the baseline designwithout blade perforation. Besides, it may move the possible tran-sient flow-separation point to further downstream, which is equiv-alent to the BLS. This reduces the size of the separation zone, thevortex strength and size; thus, the noise is also reduced. However,high perforation ratio, large hole diameter, or inappropriate perfo-ration position is likely to adversely influence or disturb thesteady-flow field and impair the aerodynamic performance.

The perforated rotors and a close-up of the perforated frontrotor are shown in Fig. 7. They are fabricated with the techniqueof rapid prototyping. The diameter of the holes d is 0.7 mm, andits perforation ratio r is about 3%. The perforation region of thefront rotor is distributed from its 85% chord to its trailing edgewhile that of the rear rotor stretches from its leading edge to its15% chord.

4.3 Separation Distance. Figure 8 shows the contours of rel-ative velocity magnitude at the section of Z¼ 2, 4, 9, and 14 mmfrom the origin located at the trailing edge of the front rotor fromthe previous computations, where Z denotes axial direction. It canbe easily found out that the wake was very strong within 9 mmfrom the trailing edge of the front rotor. Beyond this distance, thevelocity defect decayed and dissipated quickly before impingingon the downstream rotors. Note that the spanwise difference ofvelocity defect was caused by the smaller axial chord on the tipsection in the design. Besides, the tip leakage vortex was alsonoticed in Figs. 8(a)–8(d). Figure 9 shows the contours of relativevelocity magnitude at the section of r¼ 40 mm. The axial separa-tion distance between the two rotors was 30 mm and the axialchords of them were both 22 mm at this section. Since the fluid

Fig. 7 CAD pictures of perforated front rotor (top left), perfo-rated rear rotor (top right), and a close-up of the perforatedfront rotor (bottom)

Fig. 8 Contours of relative velocity magnitude (m/s) for axial positions of (a) Z 5 2 mm, (b)Z 5 4 mm, (c) Z 5 9 mm, and (d) Z 5 14 mm, Z 5 0 being the trailing edge of the front rotor

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regions of the front and rear rotors were set in two different rotat-ing reference frames as described in the Sec. 3.1, the dividingplane (also the mixing plane) was located at Z¼ 16 mm from theorigin. We can clearly see how the wakes of the front rotordecayed and dissipated downstream.

The acoustic experimental results are shown in Fig. 10, andthey indirectly validate the computational predictions. It gives themeasured noise at the fan inlet and the lateral side in the anechoicchamber for different axial spacing between the trailing edge ofthe upstream rotor and the leading edge of the downstream rotor,here denoted as S. We can see that the SPL decreased exponen-tially with S. When S was larger than 9 mm, it has little effect onthe noise level. On the contrary, the interaction noise was verysensitive to S in the range of 0–9 mm. From the spectral compari-son of the lateral noise given in Fig. 11, the most prominent peak(1567 Hz) was greatly reduced when S increased from 4 to 9 mm.In fact, the fan generated very unpleasant sharp noise, which wasrelated to the most prominent interaction frequency in its spec-trum when S¼ 4 mm. In contrast, the spectrum of S¼ 9 mm wasmore broadband. Therefore, the axial spacing of the following

study was fixed at 4 mm for the purpose of suppressing suchunpleasant interaction noise with the important gain in spacing.

4.4 Noise Comparison. Since the noise comparison would bemeaningful only under the same working condition, the aerody-namic unloading effect of the blade perforations is studied first. Inthe initial design, apertures of diameter 0.7 mm were used and fur-ther adjustment and optimization are left to future studies. Thedesign without blade perforations at the design rotating speeds isregarded as the baseline design. The characteristic curves of thebaseline design and perforated blades at the design rotating speedsare given in Fig. 12, marked with triangles and circles, respec-tively. It can be seen that there was a less than 8 Pa reduction oftotal pressure in the whole measurement range by blade perfora-tions, which was expected. The dotted curve marked with squareswas the improved characteristic curve of perforated blades afterthe voltage of the front rotor was increased by around 0.6 V, thusincreasing its rotating speed from 3500 rpm to 3600 rpm. FromFig. 12, we can see that the aerodynamic performance of the

Fig. 9 Contours of relative velocity magnitude (m/s) for r 5 40 mm, with mixing plane at 16 mmfrom the trailing edge of the front rotor

Fig. 10 Variation of SPL with axial spacing Fig. 11 Spectra comparison of the side noise

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perforated blades was improved to the same level as that of thebaseline design when the rotating speed of the front rotor wasraised to 3600 rpm.

The noise of the baseline design and the design with perforatedblades, whose axial spacing S were both 4 mm, was measured inthe anechoic chamber with a sound level meter under the condi-tion of the same aerodynamic performance but different rotatingspeeds. The sound was measured at an interval of 30 deg along acircle of 1 m in diameter on the central horizontal plane with thefan at the center. A noise time sequence of 30 s was recorded foreach direction with the sampling frequency of fs¼ 48 kHz. Forbetter comparison, the results, rendered as a polar plot of SPL indB versus angle, are shown in Fig. 13. Here, 0 deg represented thefan inlet while 180 deg denoted the outlet. We can see the two fandesigns almost had the same noise directivity and the perforatedblades had over 6–7 dB reduction in all directions.

The spectra at 330 deg for the two cases were given in Fig. 14.The first BPF of the front rotor at design rotating speed was 3500/60� 7¼ 408.3 Hz denoted as f1 while that of the rear rotor was3000/60� 5¼ 250 Hz denoted as f2. When the rotating speed ofthe front rotor was raised to 3600 rpm, its first BPF becamef1

*¼ 3600/60� 7¼ 420 Hz. The most prominent peaks for base-line design and perforated blades appeared at 2f1þ 3f2¼ 1567 Hzand 2f1

*þ 3f2¼ 1590 Hz, respectively. The SPL dropped by more

than 7 dB at the prominent peaks. It should be noted that therewas an obvious noise increase in the frequency range near15 kHz.

Figure 15 shows the noise comparison in different frequencyranges at the 330 deg position. It can be seen that the noise wasreduced mainly in the frequency range below 5 kHz through bladeperforation. The reduction magnitude was 8 dB for 0–2.5 kHz and5.6 dB for 2.5–5.0 kHz. The phenomenon of noise increase mainlyhappened in the range of 12.5–15 kHz, whose magnitude is6.8 dB. This SPL increase was related to the increased turbulentflow caused by the perforation. In other frequency ranges, theSPLs were basically unchanged.

5 Conclusions

The reported idea of perforated trailing edge for the upstreamrotor and perforated leading edge for the downstream rotor for CRfans was experimentally studied, which indicated a good result ofnoise reduction under the condition of the same aerodynamic per-formance. The design of the blade perforations was very prelimi-nary without any optimization. Nevertheless, the followingconclusions can be drawn from the initial studies:

(1) The acoustic design of perforated trailing edge for theupstream rotor and perforated leading edge for the

Fig. 12 Characteristic curves of baseline and perforatedblades

Fig. 13 Comparison of SPL directivity for baseline and perfo-rated blades

Fig. 14 Spectral comparison of baseline and perforated bladesat 330 deg

Fig. 15 SPL distribution in different frequency range

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downstream rotor had aerodynamic penalty as expected,which was a less than 8 Pa reduction of total pressure, thebaseline pressure rise being 50 Pa.

(2) The perforated blades had over 6–7 dB overall noise reduc-tion at all directions around the fan center compared withbaseline design under the same working condition and suchperforation did not change the radiated noise directivity.

(3) The spectrum at position of 330 deg indicated a most prom-inent peak at 2f1þ 3f2 and the SPL of this peak dropped bymore than 7 dB through blade perforation.

(4) The magnitude of noise reduction through blade perforationwas different in different frequency ranges. The noise wasreduced mainly in the frequency range below 5 kHz, 8 dBreduction for 0–2.5 kHz, and 5.6 dB reduction for2.5–5.0 kHz. A phenomenon of noise increase was encoun-tered mainly in the range of 12.5–15 kHz probably relatedto the increased turbulent flow through perforations.

(5) The overall SPL decreased exponentially with the separa-tion distance between the two rotors. When it was largerthan 9 mm, the axial spacing had little effect on the noiselevel.

Acknowledgment

The project was supported by a China National Key BasicResearch Scheme, or “973” scheme (2012CB7202). The firstauthor also acknowledges the support of the Ph.D. studentshipfrom the University of Hong Kong.

Nomenclature

B ¼ blade countCa ¼ axial chord

d ¼ diameter of the blade apertureD ¼ diameter of flow passagefs ¼ sampling frequency

f12 ¼ interaction frequencym ¼ spinning circumferential modePs ¼ static pressurePt ¼ inlet total pressurePv ¼ dynamic pressureQ ¼ volume flow rater ¼ given radius

Rh ¼ hub radius of bladesRt ¼ tip radius of bladesS ¼ axial spacing between two rotors

Tt ¼ inlet total temperatureDPt ¼ total pressure rise

r ¼ perforation ratioX ¼ rotating speed

Subscripts

1 ¼ parameters of front rotor2 ¼ parameters of rear rotor

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