Transmission Tonal Noise: Experimental Analysis Of The NVH
Characteristics Which Influence Vehicle Sound Quality
Jason
Jeffrey S. Williams, Project Manager
Glen C. Steyer, Technical Director
Ditman,
Project
Engineer
General Motors Corporation
General Motors Powertrain Division
Transmissions System Development Group
MC 748
Ypsilanti, MI 48198
Structural Dynamics Research Corporation
Advanced Test And Analysis Group
North American Operations
800 E. Whitcomb
Madison Heights, MI 48071
ABSTRACT
The transmission’s contribution to vehicular fonal noise is a
very significant concern for today’s automotive NVH
engineers. In general, the overall noise /eve/s generared
by
transmissions are of less concern than the structure-borne
and airborne ‘“whines” and “wh;stles” creared by the rotating
internal components. This paper explains the influence and
interaction of the source excitation with fhe m o d a l
characteristics of noise path components such as
transmissions cases, covers, brackets, mounts, propeller
shafts, suspensions, and body acoustic sens~tw~t~s.
The
implementation of practical test methods for characterizing the
excitation, component resonance effects, and the noise pafh
dynamics are presented along with data analysis techniques
designed to isolate and identify the contribution of each to a
rona/ noise annoyance.
1.0
This means that it is necessary to establish testing techniques
and test data analysis methods that characterize the total
transmission noise in a full vehicle, and break it down into the
contribution from each individual noise path. Furihermore,
it is
desirable to define each noise path contribution in terms of (1)
transmission NVH performance that can be measured on the
dynamometer or analytically predicted and (2) noise path
resonances or sensitivity effects which are specific to the rest
of the vehicle.
Additionally, the transmission NVH
performance can then be dissected into individual source
component contributions. Hence, the transmission subsystems can be developed and designed early in the design
cycle while still incorporating the NVH effects of the entire
vehicle system to predict the in-vehicle tonal noise
contribution.
These predictions can be used to set
development priorities or to design NVH characteristics which
are tuned to the vehicle platform’s specific characteristics.
INTRODUCTION
2.0
In recent years there has been a continual improvement in the
overall noise, vibration, and harshness (NVH) characteristics
of automotive systems. This has resulted in an overall
reduction in the broadband noise levels in the passenger
compartment. Simultaneously, new vehicle programs are
demanding lighter and mws fuel efficient vehicles. This often
results in vehicle bodies that are mme acoustically sensitive
and in smaller powertrains
with higher output and components
with increasing rotational speeds. This means that, in general,
the levels of harmonic excitation are increasing and the
likelihood for the interaction of these harmonics with noise path
resonances is also increasing. Consequently, with the
decrease in the broadband masking noise levels and the
increased harmonic levels, the tonal content of the interior
noise spectra has become increasingly significant to the
passenger perceived quality of automobiles. Transmissions
are one primary source of tonal noise for which the engineering
effort to design for NVH performance has become a
challenging priority.
This paper proposes that the NVH design and development of
transmissions on the computer or in the transmission
dynamometer should incorporate the effects of this entire
noise and path system, beginning with the source excitation,
including noise path resonances, and ending with the driver
detection of tonal noise.
140
BACKGROUND: TRANSMISSION
DEVELOPMENT
NVH
Generally, the transmission tonal noise, created by the
harmonics of the rotating and meshing internal components,
has a much more significant effect on the vehicle sound quality
than the transmission’s contribution to the vehicle’s overall
noise level. This tonal noise is described as a moan, whine, or
whistle that can be as much as 100 Hz wide in frequency
content and may be constant in frequency scr~ss a speed
sweep or it may vary as an order of a fundamental rotation
speed. In the past, when operating noise measurements have
shown order related noise appearing in the passenger
compartment spectrum, the trend has been to attribute the
tonal annoyance solely to the transmission design or internal
components. However, as so often is the case, it is generally
a system effect that causes the harmonics to generate a
disturbing level of interior noise. Specifically, it is comrn~n for
the interaction of the transmission excitation and the driveline
and/or the noise path resonances to generate excessive tonal
noise. This can result from the amplification due to bending or
torsional resonances in the system. Additionally, these
system resonances can have a magnification effect on the
source itself by exciting the output shaft and distorting the
alignment with gear sets, clutches, etc.
Therefore, to develop and design for transmission NVH from a
systems approach, it is necessary to isolate and characterize
these influences and then define an acceptable limit for each
parameter so that separate design groups can develop to the
acceptable limits accordingly. Transmission tonal noise can
initially be approached by consideration of these three factors:
1.
2.
3.
The most common mechanism is gear meshing. This creates a
fundamental order, as well as, sidebands. The fundamental
harmonic gear-mesh frequency for an ordinary gear is simply
the number of teeth on a gear times the rate of rotation. The
determination of gear-mesh frequencies for planetary gear
sets, usually found in automatic transmissions, is more
complicated and is explained in Reference [l]. Additionally,
vane pass frequencies of a sliding vane pump can be
calculated as the number of vane impacts times the rate of
rotation, chain and sprocket meshing can be calculated as the
number of sprocket teeth times the sprocket rotation rate, and
bearing pass frequencies can also be calculated [Z]. With this
information the NVH engineer can monitor the order related
vibration signatures and directly relate the change in these
signatures to a specific excitation source.
The source excitation or forcing function
The noise path amplification or attenuation
The ability of the passenger to detect the tonal
noise in the presence of other vehicular noise.
In general, the transmission NVH development approach has
been to attack the excitation sources and minimize the
excitation and overall noise as much as possible. There are
many references presented in the past that describe methods
for designing these components to minimize noise producing
excitation. For example, gear tooth rigidity, pitch, and profiles
can be optimized, system phasing or counter phasing can be
utilized, precision balancing can be employed, super-finished
bearings can be used, pressure fluctuations can be shaped,
isolation and damping can be introduced, and tolerances in
general can be reduced.
The transmission housings, covers, and output shafts interact
to some extent with all components and therefore must also be
evaluated for effects on vibrational inputs and the ability to
radiate airborne noise. Optimally, cases, covers, and shafts
should be designed so that the structural resonances and
radiating panel modes do not interact with excitation from the
internal rotating components. A quiet component on a bench
test can often be noisy cmce installed in the transmission due
to interaction with structural resonances. Understanding this,
the harmonic sources are generally measured at the mounting
locations on the case or cover, shaft transmitted vibration is
measured at the output shaft, and noise is measured at the
transmission exterior. This method characterizes the source
and the transmission structural effects as one measurement.
Methods to take this one step further by characterizing the
sources on a component bench test and then incorporating a
noise path effect from the source to the transmission
housings, covers, and shafts are not presented in this paper
but are under development.
However, reduction in source excitation levels is just one of
the many competing objectives in the early design stage.
Cost, durability, fuel efficiency, and packaging are other
criteria which compete for design approval. The engineer’s
freedom to design in these modifications is constrained by
these competing factors and by difficulty in proving the NVH
significance of each modification long before final system
hardware is available. Therefore, it is essential to develop
bench test and analytical methods of characterizing the level
of improvement which each modification can provide when
incorporated into the full vehicle system.
3.2 Defining The Transmission Noise Paths
Inherently, harmonic excitation will always be present at some
level due to the mechanics of the system and the inability to
manufacture precise and perfect parts. Therefore, practical
and economical limits on the degree of source attenuation that
can be achieved often exist, and it is through defining the
acceptable limits and then fine tuning the rest of the noise path
system that optimum NVH efficiency can be achieved.
The transmission noise paths of both front and rear wheel drive
vehicles generally fall into one of these three categories:
1. Elastomeric Mount Paths (structure-borne)
2. Rotating Shaft Paths (structure-borne/airborne)
3. Radiating Transmission Surfaces (airborne)
Methods are presented here for characterizing the source level
of each harmonic, the noise path amplification or attenuation
effects for that harmonic, and for predicting the resulting
interior tonal noise attributed to that harmonic. This was done
by first defining the harmonic sources and the noise paths for
transmissions and then defining the characterization methods,
and finally by defining the relationships between each and the
total passenger compaltment
noise.
For FWD transmissions specifically, several noise path
analysis studies have shown that the primary noise paths for
tonal noises are (1) structure-borne through the transmission
mounts, (2) structure-borne through the driveline half-shafts,
and (3) airborne from the radiating surface of the transmission
and across the front-of-dash barrier.
Similar studies regarding the tonal noise routes for RWD
transmissions have shown that the primary noise routes to
consider are (1) structure-borne through the transmission
cross-member mount (engine mounts are generally not a
significant noise path), (2) structure-borne propagation down
the propeller shaft and through the rear suspension, (3)
structure-borne down the propeller shaft and through the
center bearing, if applicable, (4) airborne from radiating
surfaces of the transmission across the tunnel barrier, and (5)
airborne from the propeller radiating surfaces across the tunnel
and under the carriage barrier. Figure 1 below shows a noise
path schematic of a typical RWD transmission application.
3.0 DEFINING THE HARMONIC SOURCES AND
THE POTENTIAL NOISE PATHS
In order to develop methods of experimentally characterizing
the contribution of individual noise paths and harmonic
sources. each significant path and harmonic excitation source
must first be identified and defined.
3.1 Identifying The Harmonic Sources
A transmission is the source of many sets of harmonic forces
generated by rotating components such as gear sets, vane
pumps, chains and sprockets, torque converters, bearings,
and shaft imbalances. Generally, these sources can be
identified by the harmonic vibration signature they generate.
141
it is effectively a fixed sample order track method. This
method has been chosen for this application because it allows
the NVH engineer to easily browse the spectra and analyze
each order for every measurement by quickly slicing and
comparing results.
The order slicing approach provides for acquiring an unlimited
number of slices and data channels with easy interactive
control over the slicing parameters with very little
computational overhead required. Alternative methods of
order extraction such as synchronous sampling or Kalman
filtering may also be used with some increase in extraction
accuracy, but at a significant cost in flexibility and
computational efficiency.
L
4.0
4.2 Characterization Of The Noise Paths
EXPERIMENTAL CHARACTERIZATION OF
THE CONTRIBUTORS TO TONAL NOISE
An empirical system noise model has been developed to
determine the contribution of each significant tonal noise path
to the overall tonal noise in the passenger compartment for a
given harmonic excitation source. The equation below
expresses the total driver’s right ear (DRE) tonal noise as a
function of the total tonal noise from the mount paths, the shaft
paths, and the airborne paths.
Using experimental data acquisition and analysis techniques
the source level of each harmonic can be quantified on the
transmission dynamometer. The transmission noise path NVH
performance is defined by the modal properties of the noise
path components, and therefore, artificial excitation
techniques are useful methods for evaluation. However the
methods described below go beyond just modal analysis by
describing the effect of the path modal parameter on the
interior noise due to a unit input. The methods, which are
described below have been used to establish the source
excitation and the path contribution effects for each harmonic.
To further define this equation the contribution to each type of
noise path must be defined as explained below.
4.1 Measurement Of The Harmonic Source
4.21 Elastomeric Mount Paths
The first step in characterizing the source of a transmission
tonal noise is to reproduce the operating conditions and the
noise in a controlled environment.
This is usually
accomplished on a hemi-anechoic
transmission dynamometer.
The transmission loading, speed, acceleration rate, gear
selection, and oil pressure can be accurately controlled to
simulate the real world noise in a repeatable manner. This
testing environment isolates the harmonic source and
characterizes the contribution of each source without
influence from vehicle specific characteristics.
This is the most commonly considered structure-borne path.
There have been several papers written on the subject of noise
route tracking or noise path analysis over the last ten years
which present this approach, generally for engine noise
studies [3-71. Our purpose in this section is to review the
methodology and point out a few important techniques and a
few traps to be aware of.
PDREnotal = PDnEimt + bRE/Shdt + pmwair
Using the dynamic stiffness method, a single noise route
through an elastomeric mount to a selected microphone
location in the passenger compartment can be defined as:
In general, a typical transmission whine or whistle will exist at
varying magnitudes depending on gear selection and the
engine and/or vehicle speed. Therefore, to completely
characterize a component throughout the operating envelope,
a speed sweep is generally conducted for each gear selection
and operating condition. With this approach, the standard
measurements to be made are:
(2)
(3)
where Fi(o)=Ki(w).AXi(w)
* Vibration at the mounting locations or bracket tips
* Translational and torsional vibration at the output shafts
* Noise radiating from the transmission surfaces.
From the numerous studies conducted to establish these
methods, it has been found that the measurement method of
acquiring a fine resolution spectral map for each location as a
function of the engine rpm and the output shaft rpm is the most
practical means of completely characterizing the sources.
From this measurement each order can be identified, along with
potential surprise orders, and then a post processing
technique called “slicing” can be used to extract the level of
each individual order of interest.
The slicing technique is a method of remapping the fixedsample spectral map data into the order domain, and therefore,
142
(1)
(4)
PonEim,(w)
= the total internal SPL due to the elastomeric noise paths
Pi(o) = the contribution to Ponv,, of a specific elastomeric path (i)
F!(o) = the operating force acting on the noise path (i)
+I) =
the Body Sirrcctuml
A,:ousI~(. Sensi,iviry
Func,i,,n for path (i)
K,(o) = the dynamic stiffness function of the elastomeric for path (i)
AX;(o) = the relative operating displacement across the elastomeric(i)
This expresses the noise from a single elastomeric noise route
as a contributor (Pi) to the total internal SPL from all
elastomeric routes (PDRElm,).
This contribution is calculated
by multiplying the operating force (Fi) by the Sody Structural
Acoustic Sensitivify or Transfer Function (p/f)i. The operating
force is determined by multiplying the relative displacement of
the elastomeric mount (A&) by the mounts dynamic stiffness
function (Ki).
From this path analysis, it can be seen that the dynamic
stiffness of the elastomeric mount and the p/f are the
parameters that must be measured or analytically predicted in
order to characterize these types of noise paths and predict
responses for each path based on measured inputs.
Measurement Of The P/F Function -The most accurate means
of acquiring a p/f is to measure the transfer function for each
specific path using artificial excitation. This is done by
applying a known force to the vehicle body-side of the mount in
the direction that defines that specific path and measuring the
resulting sound pressure at the passenger compartment
microphone. The force can be applied with a modal impact
hammer or a shaker. However, the hammer is the simplest and
most practical, and in general, does not introduce any
significant error. An alternative approach is to use the
reciprocity principle and measure the induced structural
acceleration per unit volumetric acceleration from an omnidirectional acoustic source placed at the interior microphone
location.
In practice, there are two approaches for defining the location
of the artificial force application for p/f measurements. The
first is to apply the force at the body rail at the mount
attachment locations. Then the flexibility of the mount bodyside structure is theoretically accounted for in the dynamic
stiffness measurement of the mount. The second approach is
to apply the force to the mount body-side structure and to
measure only the dynamic stiffness of the mounts elastomeric
component. The studies conducted on tonal noises for this
paper have shown that the dynamics of the mount body-side
structure must be accounted for in the noise path analysis.
Depending on the specific mount, the flexibility and mass of
the body-side mount structure can be a significant contributor
to the p/f in the frequency range of interest for transmissions.
This also means that the dynamic stiffness functions used to
characterize the mounts must be a function of the elastomeric
component only. This is the most practical approach because
the dynamic stiffness curves that are typically available to use
in noise route tracking applications, generally do not provide
sufficiently accurate information in this frequency range to
include the mount’s structural dynamics.
Another issue to consider is the actual location of the applied
force on the body-side bracket. The bracket should be fixture
so a force may be applied in a way that reproduces the path of
the elastomeric spring force. It may also be necessary to
measure the p/f at several attachment bolts and then calculate
an average. Studies have shown that this measurement is very
sensitive to the exact force application points.
Additionally, it is recommended that the pow&rain be removed
from the mounts and completely decoupled from the vehicle
body system before applying the force to each path. This
leaves a body sub-system which can be independently
evaluated for the response from each force and path. If the
powertrain is left attached to the mounts, it could theoretically
respond to the body-side excitation and absorb energy or
transfer energy through other paths. However, in practice, for
transmissions it has been shown that it is generally only
necessary to remove the transmission mounts if these are the
143
only paths being measured. This can be accomplished by
supporting the powertrain from below with a soft support
system while being careful not to distort the static load at the
mount locations that were not disconnected.
At each mount location a p/f should be acquired for each
orthogonal direction. Above 800 Hz the airborne path will
usually become significant, and therefore, the p/f may begin to
include measurement error. Measured correctly, the p/f
functions will represent the response of the coupled structural
behavior of the body and the passenger compartment
acoustical cavity modes.
Measurement Of The Mount Dvnamic Stiffness - The other
noise path parameter which must be experimentally determined
is the dynamic stiffness [Ki(m)] of the elastomeric component.
The mount must be tested across the frequency range of
interest in each orthogonal direction with the appropriate
preload applied and, if possible, in a temperature environment
that simulates the operating temperature. Accurate preload
simulation is vety important since most mounts are designed to
be nonlinear with large displacements. This preload is
generally determined for a given driving condition by measuring
the static displacement or by predicting the displacement
using simulation methods. In order to be consistent with the
p/f approach previously described, and to assist in acquiring
accurate data in the higher frequency ranges. it is suggested
that each mount be fixtured in a way that eliminates the
dynamics of the mount structure. By creating a rigid mount
structure, only the dynamic stiffness of the elastomeric
component will be measured.
For a given harmonic or component the tonal noise contribution
through the mount paths can be predicted using Equation (2).
4.2.2 Rotating Shaft Paths
This noise path is the most complicated to characterize. For
front wheel drive transmissions this path is purely structureborne as torsional fluctuations and translational vibrations of
the half shafts propagate to the wheels and up through the
front suspension. For the rear wheel drive applications this
path can be both structure-borne and airborne because it
includes not only torsional and translation structural
vibrations, but also the noise radiating from the propeller shaft
due to the excitation of cylinder modes. In addition, some
RWD applications implement a center bearing to couple two
propeller shafts. This creates an additional structural path to
the body.
Therefore, in general, this noise route is a function of the
torsional and translational acoustic sensitivities of the vehicle
downstream of the output shaft. A dynamic stiffness
relationship defining the transfer of forces from the output
shaft to the drive shaft is not practical in this situation because
there is not an elastomeric coupling of components.
Therefore, a mobility coupling method is used to determine the
force relationship.
The equation for this noise route is:
(5)
To calculate the franslational
PDnvshan(ro) = the total interior SPL due to the shaft noise paths
Pi(o)=
the contribution lo the PonEish,,r of one shaft path (i)
the operating angular velocity fluctuations of shaft(i)
Q(W) =
the rwsional acwrric iensiriviry.funcrion for the shaft (i)
the operating translational veIoc’Ry
of shaft (i)
These equations express the noise contribution from a single
output shaft (pi) as a function of the torsional and translational
excitation.
The torsional contribution is calculated by
multiplying the angular velocity fluctuation (ni) at the output
shaft by the shaft torsional acoustic sensitivity function (p/R),
and the translation contribution is calculated by multiplying the
translational velocity (Vi) at the output shaft by the shaft
translational acoustic sensifivify (p/v). I” addition, if a center
bearing exists, the contribution of this path is included in the
measurement of these two acoustic sensitivity functions.
Knowing these relationships, it is necessary to measure or
predict the various sensitivity functions in order to relate the
torsional and translational operating measurements from the
dynamometer to the in-vehicle scenario.
Measurement Of The PN And P/R Functions - To implement
the mobility coupling approach for the rotating shaft path it is
necessary to determine both the translational and the torsional
mobility functions and the corresponding vehicle acoustic
sensrtwrtres due to a translational or a torsional input at the
output shaft. Applying a torsional input and measuring a
rotational response can be difficult. Therefore, using the RWD
application, a practical method for determining both mobilities
from easy to measure AJF and P/F transfer functions has been
developed. This technique can also be applied to the half
shafts for FWD applications.
The transmission is
disconnected from the elastomeric mounts to decouple the
system from the body as is done in p/f testing. Then the
vehicle’s front wheels are placed up on blocks to apply a
preload to the drive shaft in order to properly represent the
operating noise path. The shaft is left attached at each end.
As shown in Figure 2, a radial impact is made with a impact
hammer on each of the clevis flanges and a driving point, a
cross frequency response function, and a p/f are measured for
=I =2 p2 The mobilities relative to
=1 =z PI & --:--;-).
each input (f:T;T
f2 f2 f2
1 1 1
the shaft center and the torsional and translational
sensitivities are all calculated from these six functions.
Drive Shaft Clevis
dtiving po;nf at the shaft center
(z), a linear superposition of load cases is used with Fl=1/2
and F2=1/2 in order to derive a transfer function for a unit
input of force at the shaft center.
&= ( 1% :’ =c )
fc
2
1,
2 f*
and a, = ;(a,++) therefore;
(7)
A frequency integration is then performed to generate the
translational mobility function
5 df = vc.
I fc
fc
(9)
Using the same superposition of forces the translat;onal
function can be calculated as show” below:
P/F
Therefore, the translational acoustic sensitivity is defined as
the pressure at the DRE for a unit velocity response at the
transmission output shaft and this is calculated as follows:
(11)
The torsional driving point can be obtained by using a linear
superposition of load cases to generate a transfer function for
a unit input of torque. Therefore, r x Fl= l/2 and r x F2= -l/Z.
a,=ra
;
+=-,a
and
.=al-a2
2r
f, =-f, = &
(12)
(13)
By substituting for a we get :
Again, using a superposition of load cases we can calculate
the pressure at the driver’s ear due to a unit of torque
fluctuation (%) or the torsional acoustic sensitivity
Then to get a torsional sensitivity in terms of a unit angular
velocity, the torsional driving point is frequency integrated
and the torsional mobiljty
function is calculated as follows:
accurately. Additionally, Reference [S] suggests a more
refined method of differentiating sound packages by
comparing a curvefit of the transmission loss slopes to the
mass law prediction.
(16)
By substituting these sensitivity functions from Equations 18
and 11 into Equation 6, the total harmonic noise from the shaft
paths can be calculated using the operating output shaft
angular and translational velocity order slices as measured in
the transmission dynamometer.
4.2.3 Airborne Paths
In order to completely account for the transmission’s tonal
noise paths, it is also necessary to characterize the airborne
noise route. The airborne contribution to tonal noise detected
in the passenger compaltment
is a function of two factors: (1)
tonal noise radiated by the transmission structural surface and
(2) the ability of the vehicle body and corresponding acoustic
barrier treatments to attenuate the noise.
The radiated noise will be measured in the transmission
dynamometer using a methodology designed to predict the
impinging noise field at the vehicle body panels. Then this
noise field will be multiplied by a function which expresses the
noise attenuation across the body. This function has been
labeled the Body (airborne) Acoustic Sensitivity Function or
P/P, sometimes also loosely referred to as the ‘transmission
ICES”.
Implementing this approach the airborne noise
contribution can be expressed as:
(19)
PDREia,,(o)
P”,(O) =
P
4.3 Establishing Allowable Limits
Using the harmonic source level and the noise path
sensitivities, the total tonal noise in the vehicle can be
predicted as described previously. The final parameter needed
to define acceptable design limits is to determine the ability of
the driver to hear the tonal noise in the presence of other
vehicular noise. Therefore, for each type of operating
condition the tonal noise can be superimposed upon the other
tonal and random noise in the vehicle to determine if it will be at
a level that is of concern. Experimental or analytical noise can
be used to simulate the other vehicle noises and Sound Quality
systems can be used to establish subjectively acceptable
levels for each type of operating condition. This limit can then
be broke down into limits for each sub-system and noise path
parameter. These limits are used to set priorities early in the
development program of a transmission.
5.0 TRANSMISSION TONAL NOISE SYNTHESIS
USING EXPERIMENTAL DATA
= the total interior SPL due to the airborne path
the radiated noise field on the transmission side of the bodv
This equation defines a relationship between the noise field at
the vehicle body barrier (P,f) and the acoustic (airborne)
sensitivity function (p/p) which is measured using artificial
excitation.
Measurement Of The PIP Functions To measure this pip
function the vehicle is typically placed in a hemi-anechoic
chamber and several speakers are arranged to produce an
artificial noise field at the front-of-dash location and/or the
tunnel and under carriage locations. The artificial noise field is
measured by placing an array of microphones at the vehicle
barrier surface. An average of these measurements is then
used as the (P,f) noise field measurement and the resulting
interior noise is also measured.
The interior noise
measurement can then be divided by the noise field average to
generate the acoustic sensitivity function (F). Depending
on the application, the characterization of this function can be
simplified by applying a mass law curvefit of the form
TL,,m,x.+20L+&j>
Using the measured or modeled pip function in Equation 19,
the airborne tonal noise contribution for a specific harmonic or
component can be predicted using the order slice noise
measurements from the transmission dynamometer. These
noise measurements should be representative of the noise
field which will be at the vehicle body barrier. Therefore, the
typical approach is measure an array of several microphones
near the transmission surfaces which are on the vehicle barrier
side of the transmission. These measurements can then be
scaled to simulate the noise field at the reflecting plane and
averaged or enveloped to predict the in-vehicle noise field.
whwzh generally predicts the slope of
the trequency dependent transmission loss (TL) function
145
With the harmonic source levels, noise path sensitivities, and
acceptable limits for each vehicle program applicable to a
specific transmission. the NVH development or design
engineer can now objectively quantify NVH oriented
modifications and focus resources on the components that will
make most significant impact in the vehicle. Additionally,
transmission NVH performance can be tuned to match
standard vehicle parameters in order to optimize the excitation
versus path sensitivity relationship.
The following example demonstrates the use of this approach
to predict the vehicle interior noise level from a pump harmonic
that was measured in the transmission dynamometer. To
simplify the prediction, each noise path was assumed to in
phase and the displacement at the body side of the
elastomeric mounts was considered insignificant compared to
the transmission side displacement. This means that the
pressures in Equation (1) become squared quantities. Using
this approach, noise and vibration measurements were made in
the transmission dynamometer in order to determine the
structure-borne and airborne source levels. In this study the
shaft paths were determined to be insignificant for the tonal
noise of interest, and therefore, they were not included in this
prediction. To predict the airborne harmonic contribution the
pump order noise slice was multiplied by the pip function. To
predict the structure-borne harmonic contribution a vector
magnitude of the vibration order was multiplied by a bracket
magnification function to predict the powertrain vector
magnitude mount displacement. Then this was multiplied by
the mount dynamic stiffness to predict the force into the body.
This force was then multiplied by an envelope of P/F functions
from all three orthogonal directions to predict the structureborne contribution. These individual contributions are
compared to the predicted noise in Figure 3.
Figure 5 - Predicted Tonal Levels ForA Full SpeedSweep Versus
An Attowabte
Limit For The Fult Speed and Frequexy
Specbvm.
6.0 SUMMARY
The complete transmission NVH design and development
program must take into account the entire noise path system.
The tools and techniques for experimentally or analytically
predicting the in-vehicle noise levels from the experimental
characterization of the noise path parameters are an integral
part of the process. The techniques and methodologies
presented here go beyond the analysis of the modal properties
by actually characterizing the effects of these modal
properties on the noise that is transmitted into the vehicle
passenger compartment. The example used shows how a total
vehicle system tonal noise concern could have been predicted
early in the transmission design and development process.
Using this prediction method, transmission components or
noise path characteristics can be tagged for further
development and the redesign results can be predicted and
compared to the baseline design.
The total predicted noise for the pump order was then
compared to the actual noise slice which was measured on a
chassis dynamometer. Figure 4 shows the accuracy of this
prediction of an established tonal noise concern. In the
frequency range from 350 - 500 Hz the prediction is well below
the in-vehicle measured levels. This suggests that the
transmission noise transmitted through the paths studied does
not dominate vehicle noise in this frequency. This prediction
would have clearly predicted the tonal noise at the level which
created a concern in the vehicle. Evaluation of the noise path
parameters versus NVH limits can direct redesign or tuning
efforts and then the new in-vehicle noise can be recalculated
to predict the tonal noise improvement.
8.0
[I]
REFERENCES
Dunlap, T.A. and H&xsen. W. G.
Tlansmisston
Noise Reductton, SAE paper no. 720735
[Z] schiltz, R.L.
For&g Frequency tdentfficatton
Ot Rdting Element Beatngs
Sound & Wbration. May ,990 Issue
[3]
Mwch, T. and W&w, P.
Noise Retie Tracting Appkations For The Transportatibn
Sound and Vibration, July 1991
[4]
Williams, R. and Salaam, M.P.
U&?Wandin
And .Wvfng Noise OuaJty
Pm. of Auto 9ech, England, Nov. ,989
151
[6]
h-Vehicle Pump Order Noise Versus
Measured In-Vehicle Pump Noise.
PraYems
Williams, J. and Stayer, G.
Experimental Noise
Aticmobites, Prcc.
Figure 4 - Predicted
,“d,,H,y
Path Anal
is For Problem Identtficatfon
of the IMA FXIII, Feb. 1995
In
Lindemann, Dr. R.. Muth, W.. Schmidt, H.
Reduction Of Transmission Noise Sy Fine Tuning Of The Drivetine
Cmpme”ts, Proc. of IMechE 1984, PaperClOB4or SAE 844010
Figure 5 shown below is an example of how the noise route
predictions can be used to predict the tonal contributions
throughout an entire speed sweep and across all relevant
transmission orders. A surface which represents an allowable
limit has been overlaid on the predicted acceleration levels in
order to evaluate the performance of the product for this
operating condition.
[7]
Storer, D. and Toniato, G.
Development Of An Expertmental
Methodology For The Anal
Structure-borne
Noise In AofonWtve Vehicles, SAE paper 95A
181
Farahanchi, F., Griffith% D.. Mason, A.S., Mayer, T.A.
Experimental Analysts Of Interior N&e Due to Powerplant
Norse. Proc. of SAE NVH Conference, ,995, SAE 95,273
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sis M
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