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1999-01-1766

Gear Noise Reduction through Transmission

Error Control and Gear Blank Dynamic Tuning

Chih-Hung (Jerry) Chung, Glen SteyerMTS Systems Corp., Noise and Vibration Division

Takeshi Abe, Mark Clapper, Chandra ShahFord Motor Co

Copyright 1998 Society of Automotive Engineers, Inc.

ABSTRACT

Gear whine can be reduced through a combination of

gear parameter selection and manufacturing process

design directed at reducing the effective transmission

error. The process of gear selection and profile

modification design is greatly facilitated through the use

of simulation tools to evaluate the details of the tooth

contact analysis through the roll angle, including the

effect of gear tooth, gear blank and shaft deflections

under load. The simulation of transmission error for a

range of gear designs under consideration was shown to

provide a 3-5 dB range in transmission error. Use of

these tools enables the designer to achieve these lowernoise limits.

An equally important concern is the dynamic mesh

stiffness and transmissibility of force from the mesh to

the bearings. Design parameters which affect these

issues will determine the sensitivity of a transmission to

a given level of transmission error. These dynamics are

studied through the use of detailed finite element models

of the transmission internals.

A systematic approach to gear element design will be

presented to optimize the gear blank design from the

perspective of the influence both the transmission error

and system dynamics on operating noise. The

correlation of model predictions with measured operating

data on prototype transmissions will be presented. The

model results will be used to illustrate how the use of

proper tuning of gear blank resonances can be used to

further reduce noise levels by 5 10 dB.

INTRODUCTION

Gear noise control measures can be categorized into the

classical areas of source path receiver measures.

Source treatments include all design and manufacturingmeasures to minimize the transmission error

Transmission error is the fundamental source of gea

whine and any reduction will result in lower perceived

levels. Tooth contact analysis incorporating mesh

kinematics, assembly tolerance analysis and load

deflections of the teeth, gear blank, and shafting enables

the engineer to minimize gear noise at the source.

The path of the gear noise in this instance is understood

to be the physical processes and system dynamics

translating the input transmission error into radiated

noise. Experience has shown that certain transmissions

are highly sensitive to transmission error, and thatproper control of the gear and shafting dynamics, as wel

as case radiation characteristics can have a dramatic

effect on reducing the transmitted noise. The system

dynamics can be properly tuned through an engineering

understanding of the underlying physical concerns and

proper design direction. This is facilitated through the

use of a detailed finite element model simulation.

Finally the receiver aspect is understood in the contex

of how the free-field radiated noise from the transmission

translates into noise inside the vehicle passenge

compartment. This includes the aspects of sound field

directivity, engine compartment reverberation, and body

panel transmission loss. The vehicle sound package is

designed to control this aspect of gear noise. The use o

engine compartment acoustic absorption and dash pane

mass damping layers is the primary design methods at

our disposal.

This paper will demonstrate how all three of the aspects

of gear whine were engineered to meet targets for a new

automatic transaxle.

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MAIN SECTION

The results presented in this paper were generated from

a combined experimental and analytical study of gear

noise in a newly developed transaxle. Overall NVH

performance was given high priority for the target

vehicle. One of the NVH goals was to minimize or

completely eliminate transmission gear whine tonaldetectability under normal driving conditions. This goal

was achieved with no slippage in program timing through

the use of vehicle level targets driven down to operating

measurements on a transmission dynamometer. Thus,

allowing for early detection of target exceedance and

resolution prior to production release.

The use of simulation technology for system dynamic

response and gear transmission error estimation

analysis was essential to clearly define the controlling

design parameters and provide direction for design

solutions. The effectiveness of design recommendations

from these simulations were confirmed throughtransmission dynamometer testing. Final testing in a

pre-production release vehicle confirmed that the target

was in fact achieved after resolving minor vehicle noise

path issues.

GEAR TRANSMISSION ERROR STUDY

One of the most significant contributors to gear whine is

gear transmission error. Transmission error is defined

as any deviation in output gear speed when the input

gear is rolled through the tooth mesh engagement with

constant angular velocity. This error can be expressed

in terms of an angular motion, or in terms of a relative

dynamic displacement along the gears line of action.

This displacement error causes radiated noise as a

result from dynamic forces at the gear tooth mesh which

are transmitted through the shafts to the transmission

housing.

Perfectly rigid gears with a perfect involute tooth profile

will theoretically have zero transmission error. In

practice, a finite level of transmission error is introduced

from a combination of gear tooth manufacturing errors,

assembly misalignments, load deflections and tooth

deformations. Some of the critical design measures forquiet gearing is the proper selection of the gear form,

refinement of the manufacturing process, and design of

the shafting and supports to minimize the resulting

transmission error.

Figure 1 shows the relative importance of gear

transmission error versus generated sound levels of fully

assembled transmissions for multiple gear sets. The

plot shows a 92% R2correlation coefficient of measured

gear transmission error versus measured transmission

operating sound levels.

-20.00

-10.00

0.00

10.00

20.00

-20 -10 0 10

Transmission Norm alized SPL (dB)

Measure

dTE

(20*LOG10(TE/Ref))

R2= 92%

Figure 1: Measured gear transmission error versus measured

transmission hemi-anechoic normalized sound pressure

levels

The transmission error was experimentally measured

with a single flank gear test rig capable of evaluating the

gear pairs under various operating loads and speeds

The sound levels shown were experimental results from

a hemi-anechoic dynamometer test system capable o

testing transmissions under load and through an

operating speed range. The normalized sound levels

shown in the plot are the de-trended average difference

over the operating frequency range relative to an

average transmission response. The two experimenta

tools presented here were important in the produc

development phase and were used to prove-ou

analytical tools and manufacturing development. Thus

they greatly improved the efficiency and timing for theNVH development of the transmission program.

Two key areas of the gear design were considered

during the development stage of the transmission. The

first consideration was the tooth macrogeometry. This

included decisions of the fundamental gear design

parameters such as module, number of teeth, pressure

angle, helix angle, etc. where careful consideration o

manufacturing, durability, and noise need to be

considered. The second phase of development was to

optimize the gear tooth microgeometry which includes

involute, lead, and bias modifications to the tooth surface

for low transmission error, optimal load distribution andoptimal contact patterns for the gear pair unde

operating loads and speed.

Transmission error minimization was one of the highest

weighted factors in the gear design process for the gea

set under study. Analytical modeling techniques and

design optimization were utilized to achieve the targe

design criteria. Gear tooth contact analysis techniques

were some of the fundamental tools which allowed the

gear engineer to optimize gear design and study gear

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stress, mesh stiffness, transmission error, load

distribution and contact patterns for differing conditions.

In order to determine the optimum tooth design a gear

design optimization software package GODA [1] (Gear

Optimization Design Analysis) establishes the basic gear

design parameters with considerations of design and

manufacturing limits. Further detail analysis is then

performed using LDP [2] (Load Distribution Program)

and CAPP [3] (Contact Analysis Program Package).CAPP is a general gear analysis tool which is capable of

solving gear tooth contact problems by combining the

strength of finite element techniques with boundary

elements and surface integral techniques. LDP is also a

contact analysis prediction tool, but uses simplified

classical representations of the tooth and gear bodies,

allowing faster detailed studies of tooth design

specifications and tooth surface topographies. Both

software techniques can determine relative transmission

error performance, mesh stiffness, stress, and tooth and

gear body load deflection. If the detailed analysis is

unsatisfactory, the design process begins again with

GODA.

-20

-10

0

10

3.5 4.5 5.5

Total Gear Contact Ratio

NormalizedSPL

(dB)

Drive Flank

Coast Flank

Trend Line

[A]

[B]

[C]

[D]

[E]

Figure 2: Normalized transmission sound pressure level (hemi-

anechoic chamber) for gear designs A-E of differing

contact ratios

Figure 2 shows five different candidate gear designs that

were considered for the transmission design.

Analytically, each of the designs had differing appeals

and compromises. With the aid of a gear tooth form

grinder all of the candidates could be prototyped at arelatively low cost and quick timing. Each design could

be easily tested in the hemi-anechoic chamber or

transmission error test stand. This allowed

determination of their relative NVH performance with

nominal tooth surface modifications, and also allowed

comparison against their analytical predictions (Figure

3).

Figure 2 clearly shows the classical trend of

transmission error decreasing with increased gear tooth

contact ratio. Gear tooth contact ratio is the theoretical

average number of gear teeth in contact during tooth

meshing. It is generally considered that smoother tooth

meshing action will occur for higher gear contact ratios

In this design range, transmission sound levels and

transmission error decrease at approximately 10 dB pe

1.0 increase in contact ratio. In this case, gear design

[C] was eventually chosen for production because it was

deemed adequate for noise performance and was

acceptable for manufacturing.

Drive Gear

Tip

Drive GearSAP

Drive GearTip

Drive Gear

SAP

Observed ContactPattern

Predicted Contact Pattern

Figure 3: LDP predicted versus observed gear tooth contact

pattern

Figure 4: CAPP FEM model of studied gear pair

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Figure 5: CAPP predicted load distribution along the gear tooth

lines of contact

Final verification of the tooth macrogeometry was

performed using CAPP. This analysis verified tooth

strength and rim deflections under load so that gear and

assembly deflections could be considered in the final

design stages. Figures 4 and 5 show examples of

CAPP simulations. Experimental verification of the

analytical tools used allow the design robustness to be

further evaluated for noise performance without

producing prototype hardware.

After the primary gear design parameters were

established, tooth microgeometry was then optimized for

the operating speeds and load deflections induced by

the transmission. Microgeometry is the refined tooth

surface topography of the gear tooth flank. Typically,

topography definition includes modifications of lead,

involute profile, and tooth bias modifications. To

determine the most robust tooth surface specifications,

optimization techniques incorporating LDP [4] were used

where manufacturing tolerances, assembly andoperating variances were considered. Table 1 shows

Taguchi allocation for five critical gear parameters: lead,

profile and bias specifications a-e, alignment and torque.

This statistical technique attempts to minimize the

transmission error over the design space considered

using historical manufacturing capability to define the

tolerance range for each parameter. This attempts to

insure that the tooth surface specification is robust

enough for transmission error over the design space.

Figure 6 shows sample order tracks of transmission

noise performance through a speed range under load

both before and after optimizing the tooth surfaces.

Table 1: Allocation to Taguchi study

a b c d e Align. Torq TE

1 + + + + + + + f1

2 + + + - - - - f2

3 + - - + + - - f3

4 + - - - - + + f4

5 - + - + - + - f5

6 - + - - + - + f6

7 - - + + - - + f7

8 - - + - + + - f8

9 0 0 0 0 0 0 0 fc

TEnominalforforceexcitingM esh-

factorweighting:5.0

)8(

indexysensitivit:2Eqn.1

minimizofunction t:1Eqn.)1(

1

2

fc

n

fin

SI

SIfcF

n

i

=

=

=

?+?=

=

This section clearly demonstrates that gear mesh

performance and transmission error are critical factors

for gear whine. Additionally, it shows how proper

analytical analysis supported by experimenta

techniques can find optimum noise performance while

considering design and manufacturing criteria. Although

gear design and manufacturing alterations can

significantly improve gear whine performance, factors

such as gear blank and transmission dynamics mus

also be considered in order to achieve total system

robustness.

Order Function

40.00

100.00

50.00

60.00

70.00

80.00

90.00

0 8002000 4000 6000Frequency (Hz)

Figure 6: Overlay of transmission sound level performance before

and after tooth profile microgeometry optimization (Solid

Optimized Profile, Dashed Initial Profile)

GEAR BLANK DYNAMICS STUDY

Unfortunately there are limits to the degree to which

transmission error can be reduced. It has been found

that certain transmission designs have a relatively high

sensitivity to transmission error. In this instance even

the best gear profile design and manufacturing controls

result in unacceptable levels of gear whine. Minimizing

the noise sensitivity to transmission error is an importan

aspect of transmission design

The design sensitivity of a transmission can be

understood in terms of the system dynamics. The

excitation is the transmission error at the gear mesh

This acts as a specified dynamic displacement forcing

the mating gear teeth apart. The gearing design mus

accommodate this transmission error, absorbing the

input motion while minimizing the vibrations transmittedto the outer casing.

Experience with previous transmission design had

shown that proper tuning of the gear and shafting

dynamics may provide on up to a 10 dB effect on

radiated noise levels. However proper tuning of the

rotating elements requires an understanding of the

component dynamics and interactions. This reqires the

use of a detailed finite element model to perform the

design studies.

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The mesh frequency for the transfer gear set in this

transmission covered the frequency range of 1500

5000 Hz under normal operating conditions. It was

necessary to use a solid element based finite element

model since the mesh frequency extended to such high

frequencies and it was anticipated that the results could

be sensitive to the fine details of the gear blank design.

Figure 7 shows the gearing finite element model. During

the course of the study a number of design iterations

were performed on the gears in order to optimize thenoise performance while not sacrificing manufacturability

nor durability. The finite element model was created

using an automatically generated mesh based off the I-

DEAS solid model. Figure 8 shows the dimensioned

wireframe used as a basis for the secondary gear model.

This modeling approach allowed for the various

wireframe dimensions to be modified, the model quickly

updated and the noise performance predicted.

Figure 7: Geometry plot of gearing finite element model.

CS1_{Global}"Revolve3"

Figure 8: Wireframe geometry basis for secondary gear blank

cross section.

Model Correlation

Model correlation tests were performed to ensure

accuracy of the predictions. Correlation was performed

using component artificial excitation data as well as ful

system operating measurements. Experimental impac

frequency response functions of acceleration over force

were compared to the model for unrestained

components such as the individual gears and shafts

Figure 9 shows a typical result from a gear blank test.

The system models were also used to predict theoperating vibration during controlled speed sweeps

This was accomplished by predicting the response pe

unit transmission error over the frequency range o

interest, then multiplying the result by the predicted leve

of transmission error from the tooth contact analysis

Figures 10 and 11 show overlays of analysis results with

experimental order track plots for the gear blank

vibration as measured through slip rings as well as

transmission housing bearing vibration. These figures

show excellent correlation at the problem frequency

range of 4200 Hz. Elsewhere the correlation was o

acceptable level for the purpose of this study.

The comparison of model predictions and test results

shown in Figures 10 and 11 provided a high level o

confidence that the model was accurately accounting fo

the relevant physical phenomena. The model was then

used to perform numerical design studies to identify an

optimal design.

0.1000

10000.0

1.000

10.00

100.00

1000.00

500 150001000 10000Frequency (Hz)

Figure 9: Overlay of an acceleration over force frequency

response function for a axial driving point on the rim of

the secondary gear. (solid test data, dashed model)

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1.000

1000.00

10.00

100.00

0 60002000 4000Frequency (Hz)

Figure 10: Overlay of model prediction with experimental data for

the order track function of secondary gear axial vibration

at the gear rim. The primary frequency of concern was

4200 Hz.

0.00

40.00

10.00

20.00

30.00

0 60002000 4000Frequency (Hz)

Figure 11: Overlay of model prediction with experimental data for

the order track function transmission housing bearing

vibration.

Design Optimization

In order to identify an optimal design it is necessary to

develop detailed insight into the physical processes and

understand the design issues. In the case of gear noise,

the optimal design is arrived at through a relatively fine

balance of the dynamic characteristics of a number of

components. An approach of blindly changing

numerous design parameters and evaluating the effect

until the magical combination of design changes is

arrived at results in a very inefficient process. However,if the baseline model is used to perform a detailed

investigation of the controlling dynamics then it becomes

possible to home in on a near optimal design in relatively

short order. The following discussion presents a quick

summary of the theory of quiet gearing design.

The dynamics of the gearing can be best understood as

a two part process of the development of dynamic force

at the mesh for a unit transmission error, followed by the

transmissibility of this mesh force to the casing. This

principle has been presented and demonstrated in

previous papers [1,2].

The fundamental design parameters which control the

dynamic mesh stiffness are gear inertias, bearing

stiffnesses and shaft bending and torsional stiffnesses

In the higher frequency ranges we find that the gear

blank out-of-plane bending modes as well as the gear

mesh compliances have a significant influence.

The transmissibility of the mesh forces to the bearings

are controlled by such parameters as the mass of the

gears, the bearing stiffnesses and the shaft bending

stiffness. Past experience of applying this simulation

technology and design optimization approach to

automotive transmissions has shown that the primary

impact of design modifications is in the control of the

dynamic mesh force. It is a common misconception in

gearing design that a low noise gear design would be

comprised of very massive and rigid gears and shafts

In fact the opposite is often true. Quiet gearing is bes

embodied in light weight gears with sufficient compliance

to absorb the transmission error without generatingundue dynamic force.

The parameters which control the dynamic mesh force

can best be understood with the aid of Figure 12. The

transmission error will be absorbed through the sum o

motion of the primary and secondary gear along the line

of action. The developed mesh force is an equal and

opposite force reaction on both gears at the pitch contac

point and oriented along the mesh line of action. Thus

the mesh compliance (or motion per unit dynamic mesh

force) is the sum of the compliance of the primary gear

and the secondary gear. The reciprocal of this mesh

compliance is then the dynamic mesh stiffness, or the

force developed at the mesh per unit transmission error.

The presence of resonances in the individual gears is

beneficial in that they reduce the mesh force

Unfortunately these resonances may also cause peaks

in the transmissibility of forces to the bearings. The

secret to the design of quiet gearing is to design in

compliances (often through modal resonances) tha

reduce the mesh dynamic stiffness but do not adversly

affect the force transmissibility from mesh zone to

bearings.

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1.00E-09

1.00E-06

1.00E-08

1.00E-07

200 80001000Frequency (Hz)

Primary N=1

N=2

SecondaryN=2

Primary N=3

SecondaryN=0

Shaft Rigid Bodyon Brgs

Mesh dynamics controlled bygear blank modes (n nodaldiameters)

F

mesh compliance

ondary primary

xf

xf

=

+ +

1

sec

_ primary

xf

secondary

xf

Figure 12: Illustration of how the mesh dynamic stiffness is the

result of the individual gear line of action (LOA) dynamiccompliances. The upper part of the figure shows the

dynamic LOA compliances for the primary and

secondary gears for the baseline design.

A bounce mode of the gear on the bearing compliances

would result in such a transmissibility peak. However a

purely torsional mode of a shaft will result in a reduction

in the mesh force with no effect on the transmissibility of

the force from the mesh to the bearings. Similarly a two

nodal diameter mode of an axi-symmetric gear blank

(the potato chipping mode) will reduce the mesh force

but will result in no net shaking force transmitted to the

bearings.

The system finite element models can be used to

accurately predict the system dynamic compliances

taking into account the gear rigid body motions and

bending compliances as well as shaft rigid body and

bending motions. These functions can be used to

understand the dynamic mesh stiffness as well as the

transmissibility of forces from the mesh to the bearings.

This process was used to study the baseline

transmission design. Figure 13 shows the dynamic

compliances and the resulting mesh dynamic stiffness.

Notice the prominent peak in the predicted mesh force atthe 4200 Hz problem frequency. This peak in the mesh

force was understood to be the result of a lack of

appropriate component modes in the neighboring

frequency range. It was determined that if certain of the

component modes could be retuned into this frequency

range then the mesh force could be smoothed out.

X/F

1.00E-09

1.00E-06

1.00E-08

1.00E-07

0 80002000 4000 6000Frequency(Hz)

A B C D

Figure 13: Gear LOA compliances and dynamic mesh force for thebaseline gear sets. Notice the irregular spacing of themodes and the wide frequency range about 4200 Hz witha total lack of component modes. This corresponds to ahigh level peak in the mesh force.

Figure 14 shows one of the gear blank modes which was

a controlling factor in the mesh compliance. This mode

is a two nodal diameter mode which occurred at 3200

Hz. A design objective was set to modify the secondary

gear web in such a manner as to force this mode into the

3800 Hz range and to drop the primary gear n=3 mode

from 5300 Hz down to 4400 Hz. This would result in a

more uniform spacing of the modes and help to reduce

the mesh force peaks in the 3000 to 5000 Hz range.

Figure 15 shows the predicted effect on the mesh

compliance and the resulting mesh force.

The actual change to the gear web wireframe was easily

developed through the use of geometric optimization

algorithm applied separately to the primary and

secondary gears. The underlying dimensions on the

wireframe were used as the design variables and the

objective functions were defined as the new targe

frequencies.

Figure 14: Deformed geometry plot of the 2ndnodal diameter mode

of the secondary gear which was tuned to the 4000 Hz

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frequency range and smooth the mesh dynamic stiffness

as a function of frequency.

X/F

1.00E-09

1.00E-06

1.00E-08

1.00E-07

0 80002000 4000 6000Frequency (Hz)

A B CD

Figure 15: Gear LOA compliances and dynamic mesh force for themodified gear sets. Notice the uniform spacing of the

modes and the alternate spacing of primary gear

resonances with secondary gear resonances. The mesh

force is seen to uniformly increase with frequency.

Prototype gears were fabricated and installed in the

transmission and operating noise measurements made

on a tranmission dynamometer. The sound pressure

order track functions were measured for four

microphones located at the front of dash locations. A

composite function was computed as the power average

of these four functions.

Figure 16 shows the predicted effect of the modifieddesign on the mesh dynamic force per unit micrometer.

Figure 17 shows the corresponding measured averaged

front of dash noise levels for the baseline and modified

designs. This figure shows data for the fundamental

gear mesh frequency as well as the second harmonic

(scaled up by a factor of 2 in order to overlay on the

fundamental).

10

70

20

30

40

50

60

700 80002000 4000 6000Frequency

(Hz)

Figure 16: Analytical results: Overlay of predicted dynamic mesh

force per unit transmission error for the baseline design

(dark dash) and the modified design with tuned gear

blanks (light dotted).

40

100

50

60

70

80

90

700 80002000 4000 6000Frequency (Hz)

Figure 17: Experimental results: Overlay of mesh order track plots

of operating noise for the baseline design (dark dash)

and the modified design with tuned gear blanks (light

dotted).

The results of Figures 16 and 17 show a remarkable

degree of aggreement in the differences between thebaseline and modified design. These both show the

modified design to be over 10 dB quieter at the previous

problem frequency of 4200 Hz, with a corresponding

increase in levels at the new resonance frequency o

3500 Hz. Also, both figures show the 3500 Hz peak to

be on the order of 8 dB lower than the original peak a

4200 Hz. This reduction corresponded to a ful

subjective rating point increase when installed in the

vehicle.

Further design modifications were studied and

implemented which affected the transmissibility of the

mesh force to the bearings. The model was additionallyused to study detail design modifications to the

transmission housing and identify stiffening structure

which resulted in 1 2 dB of additional noise reduction.

VEHICLE PATH CONSIDERATIONS

One of the difficult aspects of automotive NVH

development is the fact that numerous engineering

activities are simultaneously occuring. The subjective

NVH performance of the transmission is determined not

only by the acoustic strength of the transmission, but is

equally influenced by the acoustic integrity of the body

the presence of mechanical short circuits, and by the

ambient masking from wind and road noise. All of these

features are ever changing during the vehicle

development and do not allow for a valid subjective

evaluation of the transmission in the vehicle until the

final production release vehicle is available

Unfortunately at this stage it is too late to affect any

major reduction in the gear noise.

It is essential that a vehicle system NVH targe

development and allocation process be used. This

enables the transmission development to proceed and

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allows for rational decisions about the various design

and cost trade-offs required to have confidence in the

final product.

The dominant noise path for this transmission was

anticipated to be airborne due to the high frequencies

associated with the transfer gear. Transmission tonal

target levels were determined through a process of

applying tonal masking theory. The vehicle office had

determined targets for wind and road noise on acompetitive basis. This allows the generation of a

frequency and road speed dependent target surface for

the gear tonal noise levels at the drivers ear. The gear

whine will be undetectable for all levels below this

surface.

The drivers ear target surface is driven down to a front

of dash target surface with the use of an acoustic

transmission loss model for the dash panel. The vehicle

office again had target information for this based on

competitive analysis. The acoustic package was being

driven to meet this target.

This engine compartment front of dash noise level was

then driven down to a target level for an equivalent front

of dash plane measured on a transmission

dynamometer in a hemi-anechoic chamber. The gear

development activities identified the gear blank tuning

which enabled the design to meet this noise target. All

of the transmission noise development work was

accomplished without the benefit of full vehicle testing.

Operating road tests were performed as soon as a

production release vehicle was available. These tests

had identified a gear whine problem in the 45 mph speed

range. Subsequent testing on a chassis dynamometer

quickly determined that the noise was a result of a

mechanical path through the transmission shift cable.

Figure 18 shows an order track plot for three baseline

configuration runs and three runs with the shift cable

disconnected. The addition of mechanical isolation in

the shift cable resolved this noise problem. The gear

whine passed the subjective targets in the final vehicle

tests,

0

60

10

20

30

40

50

1000 70002000 3000 4000 5000 6000Frequency

Cable

Disconnnect

Figure 18: Passengers ear Sound Pressure order track data from

vehicle tests with and without the shift cable

disconnected.

Based on the data in Figure 18 a design modification to

the shift cable was implemented. Subsequent vehicle

drives showed the transmission noise to meet or exceed

the subjective evaluation targets.

In each vehicle development program it is essential to

perform an engineering process assessment in order tovalidate and refine the methods. Further vehicle testing

was conducted to evaluate the assumptions in the

system NVH model. The vehicle was instrumented with

four microphones in the engine compartment on the fron

of the dash. The average of the order track data from

these microphones were compared to the fou

dashplane microphone data from the transmission

dynamometer. This data confirmed the model used to

project the free-field sound pressure data to the engine

compartment data.

The apparent transmission loss was evaluated by

dividing the average front of dash microphone data by

the interior microphone data. Only the order track data

at the transfer and final drive gear mesh orders were

used. This data was compared to the anticipated body

transmission loss. These results were very close and

validated the system noise model.

CONCLUSION

The results reported in this paper demonstrate very

clearly the potential impact that simulation technology

can have on reducing transmission gear noise through

gear profile modifications for transmission erroreduction as well as gear blank tuning for reducing the

tranmission sensitivity to transmission error. The use o

tooth load deflection analysis had identified gearing

changes which resulted in a 5 dBA reduction, while

proper gear blank tuning reduced the peak noise levels

by approximately 10 dBA.

The success on this transmission has demonstrated the

value of incorporating these technologies into the early

design process. These tools will aid in optimal selection

of transmission skeletal designs and provide the ability

to develop even quieter designs before the problem has

been overly constrained.

Incorporation of these simulation technologies early in

the design process require that appropriate design

targets are developed. This begins with determination o

the target levels for vehicle interior noise based on tona

detectability theory. These targets are then rolled down

to transmission nearfield noise levels as well as moun

vibrations. Finally, these targets can be translated into

simulation targets such as bearing reaction forces and

dynamic mesh force.

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A fully integrated approach of target roll downs coupled

with simulation technologies enables the transmission

design team to select more optimal design from an NVH

perpestive at a point in the process where there exists

greater design freedom.

ACKNOWLEDGMENTS

The authors would like to acknowledge the support and

contributions of numerous colleagues at both Ford MotorCo. and MTS Systems Corp. without whom this effort

would not have been possible. Among these are Mr.

Jeff Nolff Mr. John Hiatt and Mr. Brian Wilson. The

authors are also very grateful for the solid support of

management throughout the effort and their commitment

of the necessary resources to successfully apply these

technologies.

REFERENCES

1. D. Hughson (1990), GODA5 Gear Optimization anAnalysis 5), SAE Gear Design, Manufacturing, andInspection Manual, Chap. 12, pp 175-186.

2. Houser, D.R. (1990), Gear Noise Sources and theiPrediction Using Mathematical Models, SAE GeaDesign, Manufacturing, and Inspection Manual, Chap. 16pp 213-223.

3. S. Vijayakar, A Combined Surface Integral and FiniteElement Solution for a Three Dimensional ContactProblem, International Journal for Numerical Methods inEngineering, 31, 525-546 (1991).

4. S. Sundaresan, K. Ishii, D. Houser, Design of HelicaGears with Minimum Transmission Error UndeManufacturing and Operating Variances, JSMEinternat iona l Conference on Mot ion andPowertransmissions, Nov. 1991.

5. G.C. Steyer, Influence of Gear Train Dynamics on GeaNoise, Proceedings of the National Conference on NoiseControl Engineering, pp 53-58, 1987

6. G.C.Steyer and T.C.Lim, System Dynamics in Quiet GeaDesign, Proceedings of the 9th International ModaAnalsysis Conference, pp 999-1005, 1991