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2005-01-2302

Identification and Reduction of Gear Whine Noiseof the Axle System in a Passenger Van

Sang-Kwon Lee, Sung-Kyu Go, Dongjun Yu, Joo-Young Lee and Sung-Jong KimDepartment of Mechanical Engineering, Inha University

Yun-Kyung Jo and Byung-Jae ChoNVH Research, DYMOS Co

Copyright 2005SAE International

ABSTRACT

This paper presents practical work on the reduction ofgear whine noise. In order to identify the source of thegear whine noise, transfer paths are searched andanalyzed by operational deflection shape analysis andexperimental modal analysis. It was found that gearwhine noise has an air-borne noise path instead ofstructure-borne noise path. The main sources of air-borne noise were the two global modes caused by theresonance of an axle system. These modes created avibro-acoustic noise problem. Vibro-acoustic noise canbe reduced by controlling the vibration of the noisesource. The vibration of noise source is controlled by themodification of structure to avoid the resonance or toreduce the excitation force. In the study, the excitationforce of the axle system is attenuated by changing the

tooth profile of the hypoid gear. The modification of thetooth profile yields a reduction of transmission error,which is correlated to the gear whine noise. Finally,whine noise is reduced up to 10dBA.

INTRODUCTION

In todays highly competitive automotive industry, thevehicle sound quality becomes more and more important[1-6]. There is a constant pressure to achieve a lowvehicle interior noise level in system design, especiallyfor a passenger van. For an axle system, gear whine isone of the major factors in achieving the vehicle interior

sound quality. This paper presents the results of thepractical research to reduce the whine noise of an axlesystem of a passenger van. The problem occurs in thecompartment of a passenger van. In general, gear whinenoise has two paths: airborne and structure-borne [7-8].In order to identify the transfer path of whine noise, theaccelerations on all transfer paths from the axle to thecar body are measured together with interior noise of thepassenger van simultaneously. From vibration tests, it isfound that the isolator between the body and axlesystem can reduce enough the vibration transfer energyfrom the axle system to the car body. Therefore, the

structure-borne noise is not significantly important forwhine noise. As the next process, operational shapeanalysis and experimental modal analysis are used toidentify the source of airborne noise due to the structurabehavior of the axle system during acceleration of thepassenger van. It is found that the major reason forairborne noise is the vibro-acoustic noise of the axlesystem. This vibro-acoustic noise is controlled byreducing the excitation force due to gear meshing. Theexcitation force is reduced by modifying the tooth profilewhich is related to the transmission error. Transmissionerror is very much correlated with gear whine noise [913]. In the paper, the transmission error of the ring gearhas been reduced by modification of the tooth profile ofthe ring gear under test. Finally, whine noise in thecompartment of the passenger van is reduced to 10dBAThe sound quality due to whine noise is also improved.

MEASUREMENT OF INTERIOR NOISE

An axle system with a gear ratio of 3.616 is developedThe number of teeth of a pinion gear is 13 and thenumber of teeth of a ring gear is 47 in an axle system. Apassenger van loaded with this axle system isaccelerated in order to have a subjective evaluation ogear whine noise. During acceleration, the gear whinenoise in the compartment of the passenger van isespecially heard at 60km/h and 120km/h. To identify thesource of this whine noise objectively, first, the interiornoise is measured at the front seat. Fig. 1 shows the

waterfall map for the noise signal. On this waterfall maptwo peaks are clearly identified at 1800 rpm and2300rpm as shown in Fig. 2. These two peaks occur atthe frequency band, which is the meshing frequency ofthe hypoid gear of the axle system. These peaks welcorrespond with the previous subjective evaluationMeshing frequency of the hypoid gear is 13

th times the

rotation speed of driveshaft since the number of teeth oa pinion gear is 13. In Fig. 2, it infers that gear whinenoise is related to the harmonics of the meshingfrequency of hypoid gear in an axle system.

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Speed of driveshaft (rpm)

Figure 2 Order analysis for interior noise

IDENTIFICATION OF TRANSFER PATH

Gear whine noise originates from the excitation force bythe meshing of a hypoid gear as shown in Fig. 3. Thereare two paths for whine noise: structure-borne noisepath and airborne path. The transfer of the vibrationenergy from the axle system to car body inducesstructure-borne noise. The direct transfer of vibroacoustic sound due to the shell vibration of an axlesystem from axle system to compartment of a cainduces air-borne noise.

13 order

STRUCTURE-BORNE PATH

The structural vibration of an axle system is measured aseveral points as shown in Fig. 4. Three accelerationsare measured on the housing of axle and eightaccelerations are measured on the bracket of isolationsystem during acceleration from 1800 rpm to 3500rpm odriveshaft speed. Fig. 5 shows the photo informationabout the attachment of three accelerometers attachedon the axle housing.

Frequency (Hz)

Figure 1 Waterfall map for the interior noise data

13 order

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Figure 3 Transfer paths of air-borne noise and structure-borne for

gear whine noise in an axle system; : Structure-borne

path, : air-borne path

Figure 4 Measurement point of the structure vibration of an axle

system

Excitation due to Tooth Meshing

ACCStructural VibrationOf Axle Housing ACC 3

Transfer PathResonance

ACC 2Isolation ofMount

Attenuation ofnoise by bodyBody

vibrationFigure 5 Photo explanation for measurement point of the structure

vibration of an axle systemInertial noise in the compartment of

a passenger van

Figs. 6, 7 and 8 show the waterfall map for the vibrationdata measured at the point 1, 2, and 3 of axle systemrespectively. From these results, there is one structureresonance around 3100rpm. Resonance frequency isabout 670Hz (3100rpm x 13 / 60). The other resonanceof house structure occurs at 2600 rpm. Resonancefrequency is about 1100Hz (2600rpm x 26/ 60). Two

peaks caused by resonance are clear in Fig. 6 andFig.7 , but in Fig.8 only one peak is clear caused by 13thorder of driveshaft speed. Points 1 and 2 are the placeson the hypoid gear housing. Pont 3 is the place of theaxle system. Therefore, at 670Hz, the resonance causedby 13

thorder of driveshaft speed is global mode of axle

system. At 1100Hz, the resonance caused by 26thorde

of driveshaft speed is the local mode of axle housingitself. This local mode does not contribute to the interiorsince there is no gear whine noise at 1100Hz as shownin Fig. 1. The global mode excites the whole axle systemTherefore, the axle system has high vibration level at alother four points (point 4,6,8 and 10 in Fig. 4) before the

isolator as shown in Fig. 4. It is confirmed with vibrationresult measured at point 4 as shown in Fig. 9. Theisolator, which is installed between the axle system andthe car body, attenuates this horrible vibration due toglobal mode of the axle system. Fig. 10 shows thewaterfall map for the vibration data measured at point 5This point is the place after the isolator rubber; there isno resonance peak. Although there are a fewresonances in the axle system, all vibration energy isdissipated through damping of the rubber isolator, whichis installed between the axle system and the car bodyTherefore, the gear whine noise as shown in Fig.1 is nocaused by structure-borne noise.

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Figure 7 Waterfall map for the vibration data measured at the point 2 of axle housing.

Frequency (Hz)

Figure 6 Waterfall map for the vibration data measured at the point 1 of axle housing.

Frequency (Hz)

26 order

13 order

26 order

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Figure 9 Waterfall map for the vibration data measured at the point 4 before isolation rubber

Frequency (Hz)

Figure 8 Waterfall map for the vibration data measured at the point 3 of axle housing.

Frequency (Hz)

13 order

26 order

13 order

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Frequency (Hz)

Figure 10 Waterfall map for the vibration data measured at the point 5 after isolation rubber.

Finally, Fig.11 shows the set-up of the measurementequipment for noise and vibration testing for apassenger van.

Figure 11 Set-up of measurement equipment for noise and

vibration test for a passenger van

MODAL ANALYSIS AND OPERATIONAL SHAPEANALYSIS

Although the two resonances existing in the axle system

do not affect the gear whine noise inside of the carthroughout structural vibration-transmission, they canaffect the gear whine noise as the air-borne sound. Inorder to identify the structure mode, experimental modaanalysis and operational shape analysis are usedOperational shape analysis (OSA) can be obtained byusing phase difference between reference point X andmoving point Y. The governing equation is given by [14],

)ab(j*

ii eBX

XY == (1)

where

ja

ja

eBY

eAX

=

= (2)

Fig. 12 shows the measurement point for operationashape analysis and experimental modal analysis. Thenumber of points is 125. At these 125 points, vibration ismeasured during acceleration from 1800 rpm to3200rpm of driveshaft. The phase difference for eachpoint as shown in Fig. 12 is calculated by using Eq. (1).

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Figure 12 Measurement point for operational shape analysis and

experimental modal analysis.

Operation shapes are obtained and plotted in Fig. 13.There are three major operational shapes. Two areglobal modes and the other one is a local mode. Oneglobal mode is a tensional mode as shown in Fig. 13(a),

the other a global mode is the bending mode as shownin Fig. 13(b). The local mode is the bulge mode of theaxle housing as shown in Fig. 3(c). The tensional modeof the axle system occurred at 1800 rpm with 13

thorder

of drive shaft and its frequency is 433Hz. It affects gearwhine noise and the air-borne noise. The bending modeof the axle system occurred at 3100 rpm with 13

thorder

of drive shaft and its frequency is 670Hz. It affects gearwhine noise and the air-borne noise. The bulge mode ofthe axle housing occurred at 2600 rpm with 26

thorder of

drive shaft and its frequency is 1100Hz. It does notaffect the gear whine noise as much as the two globalmodes as shown In Fig.1, since it is a local mode and

vibration energy is not enough to radiate the sound.Major three modes appearing in the operational shapeanalysis in Fig. 13 are confirmed with experimentalmodal analysis. Experimental set-up for modal analysisis arranged as shown in Fig. 14. The number ofexcitation points is the same point as points as shown inFig. 12. In order to identify the major modes of the axlesystem, the sum of transfer functions measured at the125 points as shown in Fig. 12 is calculated and plottedin Fig. 15. In this figure, two peaks at low frequencyaround 115Hz and 300Hz are found and one peak isidentified around 900Hz. The first two peaks are globalmodes: tensional mode and bending mode. The last one

is the bulge mode of the axle housing, which is a localmode as shown in Fig. 16. The natural frequency ofeach mode is shifted down to low frequency comparedwith the frequency obtained from the operation shapeanalysis as shown in Fig. 13, because the boundarycondition of the structures between operational shapeanalysis and experimental modal analysis is different.For operational shape analysis, the axle system isinstalled on the car. Therefore the structure isconstrained by the suspension system of a real car. Forexperimental modal analysis, the axle system under testis installed freely on the frame as shown in Fig. 14.Therefore, there is a frequency difference between the

resonance frequency obtained operational shapeanalysis and that obtained experimental modal analysis.

(a)

(b)

(c)

Figure 13 Operational deflection shapes during acceleration from

1800 rpm to 3200rpm of driveshaft

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(a)

(b)

(c)

Figure 14 Photo illustration for experimental modal analysis of

axle system

Figure 15 Sum of transfer functions measured at the 125 points

Figure 16 Mode shape for 3rd

mode (bulge of axle housing) of the

axle system.

DESIGN MODIFICATION OF AXLE

In the previous section it was found that structure-bornenoise does not contribute to gear whine noise becauseall vibrational energy caused by the resonance of theaxle system is damped out through the isolator between

the body and the axle system. Therefore, it is inferredthat part of the vibration energy caused by resonance ofthe axle system is radiated as vibro-acoustic energyThis vibro-acoustic energy is air-borne noise and amajor source for gear whine noise in the compartment oa passenger van. There are three methods for reducingthis air-borne noise. The first is a design modification othe car body with high acoustic transmission lossrequired. The second is a structure modification of theaxle system to avoid resonance. The last one is toattenuate the excitation force by modifying the profile othe teeth in the hypoid gear. For economical efficiency

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in the paper, the last method is adapted for the reductionof whine noise.

MODIFICATION OF TOOTH PROFILE

According to previous research results [8-13] and thecompany experience, transmission error is correlated tothe gear whine noise. In the paper, the tooth profile ismodified from a flat form to bent form listed in Table .1.

Table 1 The profile of tooth in the ring gear

In order to modify the tooth profile, the original ring gearis processed by using a 9-inch cutter, but the modified

ring gear is processed by using a 7.5- inch cutter. In thecontact area, the black area is the initial contact area ofteeth. The outline is the contact area of teeth under fullload condition. The design details of the ring gear undertest are listed in Table.2.

Table 2 The design details of the ring gear

TRANSMISSION ERROR

Gleason 500Ht is used for the measurement otransmission error of two-ring gears. The totatransmission error is reduced from 0.104 to 0.056 aslisted in Table.2. According to Table 1, the initial contactarea of two gears has the same area. However, thecontact area of the ring gear with high transmission errogoes out of the boundary of tooth face under full-loadcondition while the contact area of the ring gear with lowtransmission error is inside of the boundary of the toothface. Fig. 17 shows a photo of two ring gears. The upperone shows the ring gear with high transmission error; thebottom one shows the ring gear with low transmissionerror.Cutter

Dia.Tooth Profile Contact Ares

Original 9 Inch Slow

Modification 7.5 Inch Steep

Cutting by 9Inch Cutter

Cutting by 7,5Inch Cutter

Tooth Profile Helix Form Helix FormCutting Method Gleason GleasonCutter Diameter 9 inch 7.5 inchModule 5.268 4.907Number of Teeth 47 47Face Width 31.75 31.75

Shaft Angle 90 90

Pinion Offset 38.10 38.10

Spiral Angle 2649 225

Whole Depth 9.83 10.57Addendum 1.13 1.78Dedendum 8.70 8.79

Face Angle 7612 7612

Root Angle 749 6953

Pitch Angle 7921 75

Gear Ratio 4.89R (44X9) 4.89R (44X9)Mating Gear 3009 4580 3009 4580BacklashTransmission Error

0.15~0.200.104

0.13~0.200.056

Figure 17 Photo of two ring gears of axle system

GEAR WHINE NOISE REDUCTION

The modified axle system with new profile tooth isreplaced of noisy axle system and installed on thepassenger van under test. Interior noise for a passengervan is measured and its results are plotted in Figs. 18and 19. Vibration on the axle housing at point 2 is alsomeasured. Fig. 20 shows results of the vibration testFrom these results, high vibration around 1800rpm and3200rpm of driveshaft speed, as shown in Fig. 7, is

removed since the excitation force due to tooth meshingis significantly attenuated by reducing transmission erroalthough structure resonance exists. Therefore, Interionoise also reduced up to 10dBA at the same speed odriveshaft speed as shown in Fig. 18 and 19comparingwith Fig. 1 and2. Around 2600rpm of driveshaft speedthe 26

th order component of vibration is still high leve

since it is a local mode of axle housing and it does notsignificantly contribute to the gear whine noise.

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Frequency (Hz)

Figure 18 Waterfall map for interior noise of a passenger van with modified axle system

Figure 19 Order analysis for interior noise of a passenger van with

modified axle system

CONCLUSION

1. Basic noise and vibration tests for a passenger vanare assessed for the identification of problems.

Gear whine noise occurs around 1800rpm and3200 rpm, at which vibration level is also high.

2. Experimental modal analysis and operationadeflection shape are used to identify the vibrationtransfer path. It is found that two global modescontribute to the resonance of an axle systemaround 1800rpm and 3200rpm. The former is thetorsion mode of axle system around 433Hz; thelatter is bending mode of axle system around603Hz.Vibrations caused by these modes areattenuated through an isolator at the mount point ofthe axle system. Therefore, it is inferred that thestructure-borne noise is not contributing to gearwhine noise. However these modes contribute to

the airborne noise for gear whine noise.

3. In order to control the air-borne noise, the structureshould be modified to avoid resonance or theexcitation force should be reduced. For gear whinenoise, the modification of structure mode requiresvery high stiffness in structure. In the paper, insteadof modification of the axle system, the excitationforce is attenuated by reducing the transmissionerror by optimizing the profile of teeth of the axlegear.

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4. By reinstalling the new axle system with optimizedtooth profile on the same passenger van with gear

whine noise problem, interior noise is reduced up to10dBA.

Frequency (Hz)

Figure 20 Waterfall map for structure vibration of a axle housing measured on the positing 2 in modified axle system

ACKNOWLEDGMENTS

This work was supported by BK21 Project in Korea. Thiswork was also partly supported by DYMOS Company inKorea.

REFERENCES

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CONTACT

Dr. Sang-Kwon Lee joined Hyundai Motor ResearchCenter in Mabook-Ri, Korea, working with the

Automotive Noise and Vibration Control Group from1985 to 1994. Dr. Sang-Kwon Lee (B.E., M.E., andPh.D.) received the Ph.D. degree from the Institute of

Sound and Vibration Research (ISVR) in the Universityof Southampton, Southampton, England in 1998. He hasbeen an Associate Professor in the Department ofMechanical Engineering, Inha University, Inchon, Koreasince March 1999. His research interests are in theapplication of digital signal processing to the automotivenoise and vibration control e-mail:sangkwon@inha.ac.kr