2013 nvh str_borne_workshop_web

2013 SAE NVC Structure Borne Noise Workshop
2013 Automotive Analytics LLC A.E.Duncan
Slide 1
SAE 2013 NVH Conference
Structure Borne NVH Workshop
Alan DuncanAlan Duncan Altair Engineering @ Honda
NVH Specialist
GregGreg GoetchiusGoetchius Tesla Motors
NVH Specialist
JianminJianmin GuanGuan Altair Engineering
NVH Manager
Contact Email: aeduncan@autoanalytics.com

Slide 2
Workshop Objectives -
1. Review Basic Concepts of Automotive Structure Borne Noise.
2. Propose Generic Targets.
3. Present New Technology Example.
Intended Audience –
• New NVH Engineers.
• “Acoustics” Engineers seeking new perspective.
• “Seasoned Veterans” seeking to brush up skills.

Slide 3
• Introduction
• Low Frequency Basics
• Mid Frequency Basics
• Live Noise Attenuation Demo
• New Technology
Uncertainty and NVH Scatter
• Closing Remarks

Slide 4
• Introduction
• New Technology
• Closing Remarks

Slide 5
Ride
and
Handling
NVH Durability
Impact
CrashWorthiness
Competing Vehicle Design Disciplines

Slide 6
Automotive Engineering Objectives are Timeless

Slide 7
• Introduction
• New Technology
• Closing Remarks
Alan Duncan

Slide 8
Structure Borne Noise and Vibration
Vibrating
Source
Frequency Range: up to 1000 Hz
System Characterization
• Source of Excitation
• Transmission through Structural Paths
• “Felt” as Vibration
• “Heard” as Noise

Slide 9
Structure Borne Noise
Airborne Noise
Response
Log Frequency
“Low”
Global Stiffness
“Mid”
Local Stiffness
+
Damping
“High”
Absorption
+
Mass
+
Sealing
+
Damping
~ 150 Hz ~ 1000 Hz ~ 10,000 Hz
Automotive NVH Frequency Range

Slide 10
Low Frequency Basics
• Source-Path-Receiver Concept
• Single DOF System Vibration
• NVH Source Considerations
• Receiver Considerations
• Vibration Attenuation Strategies
Provide Improved Isolation
Mode Management
Nodal Point Mounting
Dynamic Absorbers

Slide 11
Mode Management
Dynamic Absorbers

Slide 12
RECEIVER
PATH
SOURCE
Structure Borne NVH Basics

Slide 13
Mode Management
Dynamic Absorbers

Slide 14
m
APPLIED FORCE
F = FO sin 2 π f t
k c FT
TR = FT / F
Transmitted
Force
Single Degree of Freedom Vibration
= fraction of critical damping
fn = natural frequency
f = operating frequency
( )2
n
22
n
2
2
n
ff2)ff(1
)ff(21
ζ
ζ
+−
+
=
ζ
mk

Slide 15
0 1 2 3 4 5
0
1
2
3
4
TransmissibilityRatio
1.414
Frequency Ratio (f / fn)
Vibration Isolation Principle
m
APPLIED FORCE
F = FO sin 2 π f t
k c FT
TR = FT / F
Transmitted
Force
Isolation RegionIsolation Region
1.0
0.5
0.375
0.25
0.15
0.1

Slide 16
Mode Management
Dynamic Absorbers

Slide 17
Two Main Sources
NVH Source Considerations
Suspension
Powertrain

Slide 18
Typical NVH Pathways to the Passenger
PATHS
FOR
STRUCTURE
BORNE
NVH

Slide 19
Structure Borne NVH Sources

Slide 20
Structure Borne NVH Sources
Primary Consideration:
Reduce the Source first as much as
possible because whatever enters the
structure is transmitted through multiple
paths to the receiver.
Transmission through multiple paths is
more subject to variability.

Slide 21
Mode Management
Dynamic Absorbers

Slide 22
Receiver Considerations
Subjective to Objective Conversions
Subjective NVH Ratings are typically based on a
10 Point Scale resulting from Ride Testing
A 2 ≈≈≈≈ 1/2 A 1
Represents 1.0 Rating Change
TACTILE: 50% reduction in motion
SOUND : 6.dB reduction in sound pressure level
( long standing rule of thumb )
Receiver Sensitivity is a Key Consideration

Slide 23
Mode Management
Dynamic Absorbers

Slide 24
Total 2178.2 Kg (4800LBS)
Mass Sprung 1996.7 Kg
Unsprung 181.5 Kg (8.33% of Total)
Powertrain 181.5 Kg
Tires 350.3 N/mm
KF 43.8 N/mm
KR 63.1 N /mm
Beam mass lumped on
grids like a beam
M2,3,4 =2 * M1,5
Symbolic Model of Unibody Passenger Car
8 Degrees of Freedom
From Reference 6
31
8
2
6
4
7
5

Slide 25
1 2 4 5
6 7
8
3
Tires
Wheels
Suspension
Springs
Engine Mass
Engine
Isolator
Flexible Beam for Body
8 Degree of Freedom Vehicle NVH Model

Slide 26
Force Applied to Powertrain Assembly
Forces at Powertrain could represent a First Order
Rotating Imbalance
Feng
1 2 4 5
6 7
8
3

Slide 27
Engine Isolation Example
Response at Mid Car
0.0001
0.0010
0.0100
0.1000
1.0000
5.0 10.0 15.0 20.0
Frequency Hz
Velocity(mm/sec)
Constant Force Load; F ~ A 15.9 Hz
8.5 Hz
7.0 Hz
700 Min. RPM First Order Unbalance
Operation Range of Interest
31
8
2
6
4
7
53311
88
22
66
44
77
55

Slide 28
Engine Isolation Example
Response at Mid Car
0.0001
0.0010
0.0100
0.1000
1.0000
5.0 10.0 15.0 20.0
Frequency Hz
Velocity(mm/sec)
Constant Force Load; F ~ A
700 Min. RPM First Order Unbalance
Operation Range of Interest
31
8
2
6
4
7
53311
88
22
66
44
77
55
15.9 Hz
8.5 Hz
7.0 Hz
Engine Idle Speed
Operating Shapes
at 700 RPM
Lowest Body Movement

Slide 29
Concepts for Increased Isolation
“Double” isolation is the typical strategy for
further improving isolation of a given vehicle
design.
Subframe is
Intermediate Structure
Suspension Bushing is first level
Second Level of
Isolation is at Subframe
to Body Mount

Slide 30
Removed Double Isolation Effect
Wheel
Mass
Removed
1 2 4 5
6 7
8
3

Slide 31
Double Isolation Example
Vertical Response at DOF3
0.0E+00
1.0E+00
2.0E+00
3.0E+00
4.0E+00
5.0E+00
6.0E+00
5.0 10.0 15.0 20.0
Frequency Hz
Velocity(mm/sec)
Base Model
Without Double_ISO
1.414*fn
31
8
2
6
4
7
53311
88
22
66
44
77
55

Slide 32
Mode Management
Dynamic Absorbers

Slide 33
Mode Management Chart
0 5 10 15 20 25 30 35 40 45 50
HzFirst Order Wheel/Tire Unbalance V8 Idle
Hot - Cold
EXCITATION SOURCES
Inherent Excitations (General Road Spectrum, Reciprocating Unbalance, Gas Torque, etc.)
Process Variation Excitations (Engine, Driveline, Accessory, Wheel/Tire Unbalances)
Hz
Hz
0 5 10 15 20 25 30 35 40 45 50
0 5 10 15 20 25 30 35 40 45 50
CHASSIS/POWERTRAIN MODES
Ride Modes
Powertrain Modes
Suspension Hop and Tramp Modes
Suspension Longitudinal Modes
Exhaust Modes
BODY/ACOUSTIC MODES
Body First Bending First Acoustic Mode
Steering Column First Vertical BendingBody First Torsion
(See Ref. 1)

Slide 34
Bending Mode Frequency Separation
Beam Stiffness was
adjusted to align Bending
Frequency with Suspension
Modes and then
progressively separated
back to Baseline.
1 2 4 5
6 7
8
3

Slide 35
Response at Mid Car
0.10
1.00
10.00
100.00
5 10 15 20
Frequency Hz
Velocity(mm/sec)
18.2 Hz Bending
13.Hz Bending
10.6 Bending
8 DOF Mode Separation Example
18.2 Hz
13.0 Hz
10.6 Hz
31
8
2
6
4
7
53311
88
22
66
44
77
55

Slide 36
Response at Mid Car
0.10
1.00
10.00
100.00
5 10 15 20
Frequency Hz
Velocity(mm/sec)
18.2 Hz Bending
13.Hz Bending
10.6 Bending
8 DOF Mode Separation Example
All Operating
Shapes at 10.6 Hz
31
8
2
6
4
7
53311
88
22
66
44
77
55 Highest Body Bending

Slide 37
Mode Management
Dynamic Absorbers

Slide 38
Front input forces Rear input forces
First Bending: Nodal Point Mounting Example
Mount at Nodal Point
Locate wheel centers at node points of the first bending modeshape
to prevent excitation coming from suspension input motion.

Slide 39
Passenger sits at
node point for
First Torsion.
Side View
First Torsion: Nodal Point Mounting Examples
Mount at Nodal Point
Transmission Mount of a
3 Mount N-S P/T is near
the Torsion Node.
Rear View
Engine

Slide 40
Powertrain Bending Mode Nodal Mounting
Mount system is placed to support Powertrain at the Nodal Locations of
the First order Bending Mode.
Best compromise with Plan View nodes should also be considered.
1 2 4 5
6 7
3

Slide 41
1 2 4 5
6 7
8
3
Bending Node Alignment with Wheel Centers
Redistribute Beam Masses
to move Node Points to
Align with points 2 and 4

Slide 42
Response at Mid-Car
0.0E+00
1.0E+00
2.0E+00
3.0E+00
4.0E+00
5.0 10.0 15.0 20.0
Frequency Hz
Velocity(mm/sec)
Node Shifted
Base Model
First Bending Nodal Point Alignment
31
8
2
6
4
7
53311
88
22
66
44
77
55

Slide 43
Response at Mid-Car
0.0E+00
1.0E+00
2.0E+00
3.0E+00
4.0E+00
5.0 10.0 15.0 20.0
Frequency Hz
Velocity(mm/sec)
Node Shifted
Base Model
First Bending Nodal Point Alignment
31
8
2
6
4
7
53311
88
22
66
44
77
55
Operating
Shapes at
18.2 Hz
Node Shifted Model
No Residual Body Bending

Slide 44
Diagnosis: Increase at 10.2 Hz
Body Bends up at Downward Position of Cycle

Slide 45
Motion Experienced when
Bending is Present
Motion Experienced when
Bending is Removed
Center - Undeformed Position
Up Position
Down Position
CONCLUSION:
Response Increases when a
Beneficial mode is Removed.

Slide 46
Mode Management
Dynamic Absorbers

Slide 47
YO
x
SDOF
Dynamic Absorber Concept
M
YO
x
Auxiliary Spring-Mass-Damper
m = M / 10
2DOF
M

Slide 48
Powertrain Example of Dynamic Absorber
Anti-Node Identified
at end of Powerplant
k c
Absorber attached at anti-node acting in
the Vertical and Lateral plane.
Tuning Frequency = √√√√ k/m
m
[Figure Courtesy of DaimlerChrysler Corporation]

Slide 49
Baseline Sound Level
63 Hz Dynamic Absorber
63 + 110 Hz Absorbers
Baseline Sound Level
63 Hz Dynamic Absorber
63 + 110 Hz Absorbers
[Figure Courtesy of DaimlerChrysler Corporation]
10
dB

Slide 50
Low Frequency Basics - Review
Mode Management
Dynamic Absorbers

Slide 51
• Introduction
• New Technology
• Closing Remarks
Jianmin Guan

Slide 52
Mid Frequency NVH Fundamentals
This looks familiar!
Frequency Range of Interest has changed to
150 Hz to 1000 Hz

Slide 53
Typical NVH Pathways to the Passenger
PATHS
FOR
STRUCTURE
BORNE
NVH
Noise Paths are the
same as Low
Frequency Region
Noise Paths are the
same as Low
Frequency Region

Slide 54
Mid-Frequency Analysis Character
Structure Borne Noise
Airborne Noise
Response
Log Frequency
“Low”
Global Stiffness
“Mid”
Local Stiffness
+
Damping
“High”
Absorption
+
Mass
+
Sealing
~ 150 Hz ~ 1000 Hz ~ 10,000 Hz
High modal density
and coupling in
source, path and
receiver
• Mode separation is less practical in
mid-frequency
• New Strategy is Effective Isolation:
Achieved by reducing energy transfer
locally between source and receiver at
key paths.
•• Mode separation is less practical inMode separation is less practical in
midmid--frequencyfrequency
•• New Strategy is Effective Isolation:New Strategy is Effective Isolation:
Achieved by reducing energy transferAchieved by reducing energy transfer
locally between source and receiver atlocally between source and receiver at
key paths.key paths.

Slide 55
Control Measures for Mid Frequency Concerns
Effective Isolation
Attenuation along Key Noise Paths

Slide 56
Effective Isolation

Slide 57
Classical SDOF: Rigid
Source and Receiver
TransmissibilityRatio
1.0 1.414 10.0
f / f n
“Real Structure”
Flexible (Mobile)
Source and
Receiver
Isolation Effectiveness
Effectiveness deviates from the classical development as resonances
occur in the receiver structure and in the foundation of the source.
Isolation RegionIsolation Region
1.0

Slide 58
Mobility
• Mobility is the ratio of velocity response at the excitation point on structure
where point force is applied
Mobility =
Velocity
Force
• Mobility, related to Admittance, characterizes Dynamic Stiffness of
the structure at load application point
Mobility =
Frequency * Displacement
Force
=
Frequency
Dynamic Stiffness

Slide 59
• The isolation effectiveness can be quantified by a theoretical model based on
analysis of mobilities of receiver, isolator and source
• Transmissibility ratio is used to objectively define measure of isolation
TR =
Force from source without isolator
Force from source with isolator
Isolation
V r
V ir
V is
F r
F ir
Receiver
Source
F is
Fs
V s
Isolator
V
F s=
Y i + Y r + Y s
V r
F r
Receiver
Source
F s
V s
V
F s=
Y r + Y s
V
V

Slide 60
TR =  ( Y r + Y s ) //// ( Y i + Y r + Y s ) 
• For Effective Isolation (Low TR) the Isolator
Mobility must exceed the sum of the Source and
Receiver Mobilities.
Y r : Receiver mobility
Y s : Source mobility
Y i : Isolator mobility
V m
V im
V if
F m
F im
Receiver
Source
F if
F f
V f
Isolator
TR =
Force from source without an isolator
Force from source with an isolator
Isolation
Recall that
K
1Y ∝

Slide 61
TR =  ( ) //// ( ) 
K body
1
K source
1
+
K body
1
+
K iso
1
K source
1
+
K iso
K source
K body
K iso
1.0 5.0 20.0
1.0
5.0
20.0
0.67 0.55 0.51
0.55 0.29 0.20
0.51 0.20 0.09
Generic targets:
body to bushing stiffness ratio of at least 5.0
source to bushing stiffness ratio of at least 20.0
Designing Noise Paths

Slide 62
Body-to-Bushing Stiffness Ratio
Relationship to Transmissibility
Stiffness Ratio; K body / K iso
TransmissibilityRatioTR
0
0.1
0.2
0.3
0.4
0.5
0.6
1 2 3 4 5 6 7 8 9 10
Target Min. = 5
gives TR = .20
For a source ratio of 20

Slide 63
Effective Isolation

Slide 64
Identifying Key NVH Paths
Key NVH paths are identified by Transfer Path Analysis (TPA)
Fi
Tactile
Transfer
Tactile
Transfer
Acoustic
Transfer
Acoustic
Transfer
Operating loads Operating loads
Break the system at the
points where the forces
enter the body (Receiver)
Total Acoustic Response is summation of partial responses
over all noise paths
Pt = ΣΣΣΣ paths [Pi ] = ΣΣΣΣ paths [ (P/F) i * Fi ]

Slide 65
TPA Example: Contribution at One Transfer Path
Partial response from a particular path: Pi = (P/F) i * Fi
P/T Load
Crank torque
91 Hz
FiTFi

Slide 66
TPA Example: Sum of Key Transfer Paths at One Peak
P/T Load
Crank torque
P/T Load
Crank torque
Total Response: Pt = ΣΣΣΣ paths [Pi ] = ΣΣΣΣ paths [ (P/F) i * Fi ]

Slide 67
Attenuating Key NVH Paths
TPA Example: Identifying Root Cause of Dominant Paths
P/T Load
Crank torque
PMi
Fi
TFi

Slide 68
Attenuating Key NVH Paths
TPA Example: Dominant Paths over Frequencies
P/T Load
Crank torque
Fi
TFi
PMi

Slide 69
TPA Example: Cascading Vehicle Targets to Subsystems
Limit TF to a 55 dB target
See changes in response
P/T Load
Crank torque
TFi
Once the dominant noise paths and
root cause have been identified, the
task is reduced to solving problems of:
1. High Force
2. High Transfer Function
3. High Point Mobility

Slide 70
Fi
Acoustic Transfer (P/F)i
Acoustic Transfer (P/F)i
Operating loads create
Forces (Fi) into body at
All noise paths
F F
F F
F
F F
Pt = ΣΣΣΣ paths [Pi ] = ΣΣΣΣ paths [ Fi * (P/F) i ]
= ΣΣΣΣ paths [ Fi * (P/V) i * (V/F)i]
P/V
P/F
(Kbody)
V/F
Measurement Parameters Generic Targets
P/F Acoustic Sensitivity 50 - 60 dBL/N
V/F Structural Point Mobility
(Receiver Side)
0.2 to 0.3
mm/sec/N

Slide 71
Recall for Acoustic Response Pt
Pt = ΣΣΣΣ paths [Pi ] = ΣΣΣΣ paths [ Fi * (P/V) i * (V/F)i]
“Downstream” Effects: Body Panels
(P/V)i !!!! “Downstream” (Body Panel) System Dynamics: Three Main Effects:
Increased Damping
Increased Stiffness
1. Panel Damping
2. Panel Stiffness
3. Panel Acoustic Contribution

Slide 72
Generic Noise Path Targets
K iso
K body
> 5.0
K iso
K source
> 20.0
Acoustic
Sensitivity
<
50 - 60
dBL/N
Structural
Mobility < 0.2 to 0.3 mm/sec/N
Panel Damping Loss Factor
> .10
Primary: Minimize the Source Force
< 1.0 N

Slide 73
Final Remarks on Mid Frequency Analysis
• Effective isolation at dominant noise paths is critical• Effective isolation at dominant noise paths is critical
• Reduced mobilities at body & source and softened
bushing are key for effective isolation
• Reduced mobilities at body & source and softened
bushing are key for effective isolation
• Other means of dealing with high levels of response
(Tuned dampers, damping treatments, isolator
placement at nodal locations) are also effective
• Other means of dealing with high levels of response
(Tuned dampers, damping treatments, isolator
placement at nodal locations) are also effective
• Mode Separation remains a valid strategy as modes
in the source structure start to participate
• Mode Separation remains a valid strategy as modes
in the source structure start to participate

Slide 74
Structure Borne NVH: Concepts Summary
• Source-Path-Receiver as a system
1. Reduce Source
2. Rank and Manage Paths
3. Consider Subjective Response
• Effective Isolation
• Mode Management
• Nodal Point Placement
• Attachment Stiffness
• “Downstream” (Body Panel) Considerations
• Source-Path-Receiver as a system
1. Reduce Source
2. Rank and Manage Paths
3. Consider Subjective Response
• Effective Isolation
• Mode Management
• Nodal Point Placement
• Attachment Stiffness
• “Downstream” (Body Panel) Considerations

Slide 75
• Introduction
• New Technology
• Closing Remarks
Greg Goetchius

Toolbox Demo Noise Test Results
85
61
68
63
55
47
39
30
40
50
60
70
80
90
1) Baseline:
Imbalance, No
Isolation
2) Imbalance +
Isolation
3) No Imbalance,
No Isolation
4) No Imbalance,
No Isolation +
Damping
5) No Imbalance +
Isolation +
Damping
6) #5 + Absorption 7) #6 + Insulator
Mat
SPL(dBA)Tool Box Demo Test Results

Slide 77
• Introduction
• New Technology
NVH Scatter and Uncertainty
Closing Remarks Alan Duncan

Slide 78
• Scatter 20 Years Ago
• New Technology to Address Scatter

Slide 79

Slide 80
Frequency ( Hz )
Mag.ofFRF
Magnitude of 99 Structure Borne Noise Transfer Functions for
Rodeo’s at the Driver Microphone
Measurements from Kompella and Bernhard
( Ref. 8 ) 1993 Society of Automotive Engineers, Inc.

Slide 81
SoundPressure[dB(lin)]
Frequency ( Hz )
Acoustic Scatter from Simulation of the
vibroacoustic behavior of a vehicle due
to possible tolerances in the component
area and in the production process.
© 2000 Society of Automotive Engineers, Inc.
Reproduced with permission from paper
by Freymann, et. Al. (Ref. 9)
12 dB Variation
Experimentally detected Scatter in low
frequency vibroacoustic behavior of
production vehicles.
Freymann, BMW NVH Scatter Results

Slide 82
CONCLUSIONs: From K/B and BMW Studies
It is not highly probable:
1. that a Single Test will represent the Mean response
2. that a CAE Simulation will match a Single Test
Variability observed from multiple tests of “identical” vehicles is
important in understanding the degree to which the Test (or
Simulation) of a Design is representative of the Mean response.
How many Tests (or Simulations) of a Design would be required
for the result to be considered statistically significant?
Scatter Implications for Test (or Simulation) NVH Development
Scatter is the “Physics”
Test Simulation

Slide 83

Slide 84
Test Variation Band
10. dB; 50-150 Hz
20. dB; 150-500 Hz
Reference Baseline Confidence Criterion
For Operating Response Simulations
Test Upper Bound
Test Band Average
Test Lower Bound
REF. 8
Confidence Criterion:
Simulation result must
fall within the band of
test variation.
Simulation Prediction
FUDGE
FACTORS
SoundFRF
Frequency Hz
2003 Workshop Confidence Criterion Lecture with Voice Track
available at www.AutoAnalytics.com on DOWNLOAD page.

Slide 85
FlashBack: 1995 Paper on Root Cause of Scatter
(Ref. 15)

Slide 86
Root Cause of Scatter : Conditions
Total Response

Slide 87

Slide 88(Ref. 12)
Non-Parametric Probabilistic Simulation; Soize, Durand, Gagliardini
Goal: Define a Modeling Methodology that:
1. Accounts for NVH Scatter
2. Quantifies Uncertainty with Statistical
Significance using a Stochastic Model
3. Accounts for the Combined Effect of Modeling
and Manufacturing Uncertainty

Slide 89
(See Ref. 11 for Model Details)
Tactile Responses
POWERTRAIN LOADS
NTF @ DRIVER EAR

Slide 90
Standard Eqn. for
Structural-Acoustic
Coupling Modal Model
Typical Derivation for
creating the randomized
Dynamic Matrix with
Gaussian Distribution.
Non-Parametric: Direct
Changes in Modal Matrices.
The scatter created is a
function of:
Dispersion Parameter: δ
Once determined, δ is a
constant controlling the
amplitude level of scatter.
7 Dispersion Constants
are needed:
3 for Structure: M, D, K
3 for Acoustics: M, D, K
1 for Coupling: Cn,m

Slide 91
Proof of Process Convergence
CONCLUSIONS: Converged at 200 REALIZATIONS
and with Struct to 400 and Fluid to 350 Hz
Monte Carlo
Randomization
Converges at 200
Realizations of the
Random Matrices.
Increasing No. of
Modes

Slide 92
Structural Uncertainty
Coupling
Uncertainty
Acoustic Uncertainty
95% Upper Confidence Band
Nominal / Original Model Response
50% Confidence Band (Mean Stochastic Model)
95% Lower Confidence Band
CONCLUSION:
• The Model shows 95%
Confidence Bands with
increasing Scatter at higher
frequency similar to K-B Study.

Slide 93(Ref. 13)
Main Paper - Development is Extensive
Includes Test Data Comparison

Slide 94
FIG. 10. Finite element mesh of the computational structural acoustic model.
(See Ref. 11 for Model Details)
Full Vehicle System – Body and Chassis
Structural-Acoustic Model
Similar Detail Level as first paper

Slide 95
FIG. 16. Comparisons of the stochastic computational model results with the experiments. Graphs of the root
mean square of the acoustic pressures averaged in the cavity in dB scale: experiments for the 30 configurations
(gray lines); Mean computational model (dashed line); mean value of the random response (mid thin solid line);
95% confidence region: the upper and lower envelopes are the upper and lower thick solid lines.
95% UPR
Nominal
50% Mean
95% LWR
NOTE: Acoustic Dispersion Parameters are determined here with Mean Structure held Invariant.

Slide 96
FIG. 19. Comparisons of the stochastic computational model results with the experiments for observation
Obs6. Graphs of the moduli of the FRFs in dB scale: experiments for the 20 cars (gray lines); Mean
computational model (dashed line); mean value of the random response (mid thin solid line); confidence
region: the upper and lower envelopes are the upper and lower thick solid lines.
OBS # 6
P/T Load
2nd Order
95% UPR
Nominal
50% Mean
95% LWR
NOTE: Structure Dispersion Parameters are determined here with Mean Acoustic held Invariant.

Slide 97
FIG. 20. Comparisons of the stochastic computational model results with the experiments for the booming
noise. Graphs of the moduli of the FRFs in dB(A) scale: experiments for the 20 cars (thin gray lines); mean
value of the experiments (thick gray line); Mean computational model (dashed line); Mean value of the
random response (mid thin solid line); 95% confidence region: the upper and lower envelopes are the upper
and lower thick solid lines.
P/T Load
2nd Order
95% Upr
Nominal
50% Mean
95% Lwr

Slide 98
FIG. 20. Comparisons of the stochastic computational model results with the experiments for the booming
noise. Graphs of the moduli of the FRFs in dB(A) scale: experiments for the 20 cars (thin gray lines); mean
value of the experiments (thick gray line); Mean computational model (dashed line); Mean value of the
random response (mid thin solid line); 95% confidence region: the upper and lower envelopes are the upper
and lower thick solid lines.
P/T Load
2nd Order
95% Upr
Nominal
50% Mean
95% Lwr
CONCLUSIONS:
. The 95% Confidence Bands encapsulate the Measured Scatter of 20 Vehicles.
. Half-Bandwidth Scatter (between Mean and Upr 95%) is similar to K-B Half-Bands.
NOTE: The Method has Quantified the Model and Manufacturing Combined Uncertainty.

Slide 99
(Ref. 14)
Non-Parametric Probabilistic Simulation; Jund, Guillaume, Gagliardini

Slide 100
FIG. 4: Computed booming noise inside a car vs. rpm, using only structure-borne excitation compared with the
measured value (brown bold). Thin dotted black: deterministic computation; green: median value; thin dotted
blue: lower bound with a 95% probability; dotted red: upper bound with a 95% probability.
P/T Load
2nd Order
95% Upr
Nominal
50% Mean
95% Lwr
Test
Measurement
Author:
Focus of Correlation Effort
All Test Peaks in Band: Results imply a
Countermeasure is needed.

Slide 101
FIG. 7: Booming noise inside a car vs rpm.
Stochastic modelling. Comparison of the median
value of a baseline configuration (green) with a
modification set (blue bold)
P/T Load
2nd Order
FIG. 6: Booming noise inside a car vs rpm.
Deterministic models. Comparison of a baseline
configuration (green) with a modification set (blue
bold).
DETERMISTIC MODEL STOCHASTIC MODEL - Mean Level
. Amplitude Range = 30 dB
. Need Tradeoff Judgment
. Amplitude Range = 15 dB
. Decision is Clearer that
both Designs are Equal
Performers
Blue: New Design
Green: Base
A-B Design Decision using Deterministic Model - vs - Mean Stochastic Model

Slide 102
Conclusions:
Kompella and Bernhard Test Observations are still relevant after 20 Years.
Scatter-like NVH Variation exists even in Best-in-Class vehicles.
Observations about Scatter:
Observations from Soize, Durand, Gagliardini, et. al. Papers
The Non-Parametric stochastic computational model with dispersion
parameters derived from a test database accounts for NVH Scatter due
to combined Modeling and Manufacturing Uncertainties.
The Upper, Lower, and Mean Confidence Probabilities lead to more
precise assessment of the effects of NVH scatter.
A database of dispersion parameters enables a virtual product
development process assuring robust NVH performance.
The model configuration lends itself to an automated computational
methodology driving a robust virtual product development process.

Slide 103
• Introduction
• New Technology
• Closing Remarks: Q & A
Alan-Greg-Jimi

Slide 104
SAE 2013 NVH Conference
Thank You for Your Time!
Q & AQ & A

Slide 105
Primary References (Workshop Basis: 4 Papers)
1. A. E. Duncan, et. al., “Understanding NVH Basics”, IBEC, 1996
2. A. E. Duncan, et. al., “MSC/NVH_Manager Helps Chrysler Make
Quieter Vibration-free Vehicles”, Chrysler PR Article, March 1998.
3. B. Dong, et. al., “Process to Achieve NVH Goals: Subsystem
Targets via ‘Digital Prototype’ Simulations”, SAE 1999-01-1692,
NVH Conference Proceedings, May 1999.
4. S. D. Gogate, et. al., “’Digital Prototype’ Simulations to Achieve
Vehicle Level NVH Targets in the Presence of Uncertainties’”,
SAE 2001-01-1529, NVH Conference Proceedings, May 2001
Structure Borne NVH References
WS + Refs. at www.AutoAnalytics.com at download link
Structure Borne NVH Workshop - on Internet
At SAE www.sae.org/events/nvc/specialevents.htm

Slide 106
Supplemental Reference Recommendations
5. T.D. Gillespie, Fundamentals of Vehicle Dynamics, SAE 1992
(Also see SAE Video Lectures Series, same topic and author)
6. D. E. Cole, Elementary Vehicle Dynamics, Dept. of Mechanical
Engineering, University of Michigan, Ann Arbor, Michigan, Sept.
1972
7. J. Y. Wong, Theory of Ground Vehicles, John Wiley & Sons, New
York, 1978
8. N. Takata, et.al. (1986), “An Analysis of Ride Harshness” Int.
Journal of Vehicle Design, Special Issue on Vehicle Safety, pp.
291-303.
9. T. Ushijima, et.al. “Objective Harshness Evaluation” SAE Paper
No. 951374, (1995).
10. G. Goetchius; “The Seven Immutable Laws of CAE/Test
Correlation” Sound and Vibration Mag. Editorial June 2007, online
at SandV.com

Slide 107
New Technology References (2013 NVH Workshop)
11. Sol, A.; Van Herpe, F.; “Numerical Prediction of a Whole Car Vibro-Acoustic
Behavior at Low Frequencies”; SAE # 2001-01-1521
12. Durand, J.; Gagliardini, L.; Soize, C.; “Nonparametric Modeling of the
Variability of Vehicle Vibroacoustic Behavior”; SAE # 2005-01-2385; SAE 2005
Noise and Vibration Conference Proceedings.
13. Durand, J.; Gagliardini, L.; Soize, C.; “Structural-acoustic modeling of
automotive vehicles in presence of uncertainties and experimental identification
and validation”; Journal of the Acoustical Society of America 24, p 1513-1525
14. Jund, A.; Gagliardini, L.; et.al.; “AN INDUSTRIAL IMPLEMENTATION OF
NON-PARAMETRIC STOCHASTIC MODELLING OF VEHICLE
VIBROACOUSTIC RESPONSE” CONVEBONOV Workshop, University of
Sussex, Brighton, UK, March 27-28 2012
15. Gardhagen, B. and Plunt, J., “Variation of Vehicle NVH Properties due to
Component Eigenfrequency Shifting - Basic Limits of Predictability” SAE 951302
May 1995 NVH Conference

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