2011 SAE Structure Borne NVH Workshop Web

SAE 2011 NVH Conference
Structure Borne NVH Workshop
Presenters:

Alan Duncan Altair Engineering @ Honda
Contact Email: aeduncan@autoanalytics.com
Web: www.autoanalytics.com

Greg Goetchius Tesla Motors

Kiran Govindswamy FEV Inc.

Jianmin Guan Altair Engineering
Slide 1
2011 SAE Structure Borne NVH Workshop www.autoanalytics.com

Workshop Objectives -
1. Review Basic Concepts of Automotive Structure Borne Noise.
2. Propose Generic Targets.
3. Present Real World Application Example.

Intended Audience –
• New NVH Engineers.
• “Acoustics” Engineers seeking new perspective.
• “Seasoned Veterans” seeking to brush up skills.

Slide 2

• Introduction
• Low Frequency Basics
• Mid Frequency Basics
• Live Noise Attenuation Demo
• Case Studies
• Interior Noise Diagnosis using VINS
• NVH Principles: Applications
• Closing Remarks
Slide 3

• Introduction Greg Goetchius
• Case Studies
• Closing Remarks
Slide 4

• Introduction
• Case Studies
• Closing Remarks
Slide 5

Competing Vehicle Design Disciplines

Ride Impact
and CrashWorthiness
Handling

NVH Durability

Slide 6

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 7

Automotive NVH Frequency Range

Structure Borne Noise
Airborne Noise
Response

Absorption
+
Mass
+
Local Stiffness Sealing
+ +
Global Stiffness Damping Damping
“Low” “Mid” “High”

~ 150 Hz ~ 1000 Hz ~ 10,000 Hz

Log Frequency Slide 8

• Introduction
• Case Studies
• Vehicle Interior Noise Simulation
• Closing Remarks
Slide 9

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 10

Mode Management
Dynamic Absorbers
Slide 11

Structure Borne NVH Basics

RECEIVER

PATH

SOURCE
Slide 12

Mode Management
Dynamic Absorbers
Slide 13

Single Degree of Freedom Vibration
APPLIED FORCE

F = FO sin 2 π f t

1 + (2ζ f fn ) 2
TR = FT / F =
(1 − f fn ) + (2ζ f fn )
2 2 2 2
m
Transmitted
Force
k c FT ζ = fraction of critical damping
fn = natural frequency k m
f = operating frequency
Slide 14

Vibration Isolation Principle
4
0.1
APPLIED FORCE
Transmissibility Ratio

0.15

F = FO sin 2 π f t
3

m TR = FT / F
0.25

2 k c FT Transmitted
0.375 Force
0.5
1.0
1
Isolation Region
Isolation Region

0
0 1 1.414 2 3 4 5
Frequency Ratio (f / fn) Slide 15

Mode Management
Dynamic Absorbers
Slide 16

NVH Source Considerations

Suspension
Powertrain

Two Main Sources
Slide 17

Typical NVH Pathways to the Passenger

PATHS
FOR
STRUCTURE
BORNE
NVH
Slide 18

Structure Borne NVH Sources

Slide 19

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 20

Mode Management
Dynamic Absorbers
Slide 21

Receiver Considerations
Subjective to Objective Conversions
Subjective NVH Ratings are typically based on a
10 Point Scale resulting from Ride Testing
Receiver Sensitivity is a Key Consideration

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 ) Slide 22

Mode Management
Dynamic Absorbers
Slide 23

Symbolic Model of Unibody Passenger Car
8 Degrees of Freedom

2 4
8
1 3 5
6 7

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

8 Degree of Freedom Vehicle NVH Model
Engine Mass
8
Engine Flexible Beam for Body
Isolator
1 2 3 4 5

Suspension
Springs
6 7
Wheels
Tires

Slide 25

Force Applied to Powertrain Assembly
Feng
8

1 2 3 4 5

6 7

Forces at Powertrain could represent a First Order
Rotating Imbalance
Slide 26

Engine Isolation Example
Response at M id Car
1.0000

Constant Force Load; F~A 15.9 Hz
8.5 Hz
Velocity (mm/sec)

0.1000
7.0 Hz

0.0100

0.0010
2 4
8
1 3 5
6 7

700 Min. RPM First Order Unbalance
0.0001 Operation Range of Interest
5.0 10.0 15.0 20.0
Frequency Hz

Slide 27

Concepts for Increased Isolation
“Double” isolation is the typical strategy for
further improving isolation of a given vehicle
design.

Second Level of
Isolation is at Subframe
to Body Mount

Subframe is
Intermediate Structure

Suspension Bushing is first level

Slide 28

Removed Double Isolation Effect

8

1 2 3 4 5

Wheel
6 Mass 7

Removed

Slide 29

Double Isolation Example
Vertical Response at DOF3
6.0E+00

Base Model
5.0E+00
(mm/sec)

Without Double_ISO

4.0E+00
2 4
8
1 3 5
6 7

3.0E+00
Velocity

1.414*fn
2.0E+00

1.0E+00

0.0E+00
5.0 10.0 15.0 20.0
Frequency Hz
Slide 30

Mode Management
Dynamic Absorbers
Slide 31

Mode Management Chart
EXCITATION SOURCES
Inherent Excitations (General Road Spectrum, Reciprocating Unbalance, Gas Torque, etc.)
Process Variation Excitations (Engine, Driveline, Accessory, Wheel/Tire Unbalances)
0 5 10 15 20 25 30 35 40 45 50
First Order Wheel/Tire Unbalance Hz
V8 Idle
Hot - Cold

CHASSIS/POWERTRAIN MODES
Suspension Hop and Tramp Modes
Ride Modes Suspension Longitudinal Modes
Powertrain Modes Exhaust Modes
0 5 10 15 20 25 30 35 40 45 50
Hz

BODY/ACOUSTIC MODES
Body First Torsion Steering Column First Vertical Bending
Body First Bending First Acoustic Mode

0 5 10 15 20 25 30 35 40 45 50
Hz
(See Ref. 1) Slide 32

Bending Mode Frequency Separation

8

1 2 3 4 5

Beam Stiffness was
adjusted to align Bending
Frequency with Suspension
6 7
Modes and then
progressively separated
back to Baseline.

Slide 33

8 DOF Mode Separation Example
Response at Mid Car 18.2 Hz Bending
13.Hz Bending
100.00
10.6 Bending

2 4
8
1 3 5
Velocity (mm/sec)

6 7 10.6 Hz

10.00
13.0 Hz
18.2 Hz

1.00

0.10

5 10 15 20
Frequency Hz
Slide 34

Mode Management
Dynamic Absorbers
Slide 35

Mount at Nodal Point
First Bending: Nodal Point Mounting Example

Front input forces Rear input forces

Locate wheel centers at node points of the first bending modeshape
to prevent excitation coming from suspension input motion.
Slide 36

Mount at Nodal Point
First Torsion: Nodal Point Mounting Examples

Side View
Rear View
Passenger sits at Engine
node point for
First Torsion.

Transmission Mount of a
3 Mount N-S P/T is near
the Torsion Node. Slide 37

Powertrain Bending Mode Nodal Mounting

1 2 3 4 5

6 7

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.

Slide 38

Bending Node Alignment with Wheel Centers
Redistribute Beam Masses
8 to move Node Points to
Align with points 2 and 4

1 2 3 4 5

6 7

Slide 39

First Bending Nodal Point Alignment
Response at Mid-Car
4.0E+00

Node Shifted
4
Base Model 8
2
(m m /sec)

1 3 5
6 7

3.0E+00

2.0E+00
Velocity

1.0E+00

0.0E+00

5.0 10.0 15.0 20.0
Frequency Hz
Slide 40

Mode Management
Dynamic Absorbers
Slide 41

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

Slide 42

Powertrain Example of Dynamic Absorber

Anti-Node Identified
at end of Powerplant

k c
m
Absorber attached at anti-node acting in
the Vertical and Lateral plane.

Tuning Frequency = √ k/m
Slide 43
[Figure Courtesy of DaimlerChrysler Corporation]

Baseline Sound Level
Baseline Sound Level
63 Hz Dynamic Absorber
63 Hz Dynamic Absorber
63 + 110 Hz Absorbers
63 + 110 Hz Absorbers

10
dB

Slide 44
[Figure Courtesy of DaimlerChrysler Corporation]

Low Frequency Basics - Review
Mode Management
Dynamic Absorbers
Slide 45

• Introduction
• Mid Frequency Basics Alan Duncan
• Case Studies
• Closing Remarks
Slide 46

Mid Frequency NVH Fundamentals

This looks familiar!
Frequency Range of Interest has changed to
150 Hz to 1000 Hz Slide 47

Typical NVH Pathways to the Passenger

PATHS
FOR
STRUCTURE
Noise Paths are the
BORNE
Noise Paths are the
same as Low
same as Low NVH
Frequency Region
Frequency Region Slide 48

Mid-Frequency Analysis Character
Structure Borne Noise
Airborne Noise

• Mode separation is less practical in
• Mode separation is less practical in
High modal density mid-frequency
mid-frequency
and coupling in
source, path and
• New Strategy is Effective Isolation:
• New Strategy is Effective Isolation:
receiver
Achieved by reducing energy transfer
Achieved by reducing energy transfer
locally between source and receiver at
locally between source and receiver at
key paths.
key paths.
Response

Absorption
+
Local Stiffness Mass
+ +
Global Stiffness Damping Sealing
“Low” “Mid” “High”

~ 150 Hz ~ 1000 Hz
Log Frequency ~ 10,000 Hz

Slide 49


Control Measures for Mid Frequency Concerns

Effective Isolation

Attenuation along Key Noise Paths

Slide 50



Effective Isolation


Slide 51

Isolation Effectiveness
Classical SDOF: Rigid
Source and Receiver

“Real Structure”
1.0 Flexible (Mobile)
Source and
Transmissibility Ratio

Receiver

Isolation Region
Isolation Region f/fn

1.0 1.414 10.0
Effectiveness deviates from the classical development as resonances
occur in the receiver structure and in the foundation of the source.
Slide 52

Mobility
• Mobility is the ratio of velocity response at the excitation point on structure
to the force applied in the direction of the velocity

Velocity
Mobility =
Force

• Mobility, also called Admittance, characterizes Dynamic Stiffness of
the structure at load application point

Frequency * Displacement
Mobility =
Force

Frequency
=
Dynamic Stiffness

Slide 53

Isolation
• The isolation effectiveness can be quantified by a theoretical model based on
analysis of the SOURCE - PATH - RECEIVER transfer across an isolator

• Transmissibility Ratio (TR) is used to objectively define measure of isolation
Force from source with isolator
TR =
Force from source without isolator

V Receiver Vr
V Receiver Vr
V Fr
Fr Fs =
Yi + Y r + Y s F ir
V ir
Fs
Vs Isolator V is
Source
F is
V
F s= Fs
Yr +Ys Vs
Source
Slide 54

Isolation

Force from source with an isolator Receiver Vm
TR =
Force from source without an isolator
Fm

TR =  ( Y r + Y s ) / ( Y i + Y r + Y s )  F im
V im
Isolator V if
Y r : Receiver mobility
F if
Y i : Isolator mobility
Y s : Source mobility Ff
Vf
Source

• For Effective Isolation (Low TR) the Isolator
Mobility must exceed the sum of the Source and
Receiver Mobilities.
Recall that Y ∝1
K Slide 55

Designing Noise Paths
1 1 1 1 1
TR =  ( + ) / ( + + )
K body K source K body K iso K source
K source
K body K iso
K iso 1.0 5.0 20.0

1.0 0.67 0.55 0.51

5.0 0.55 0.29 0.20

20.0 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
Slide 56

Body-to-Bushing Stiffness Ratio
Relationship to Transmissibility

0.6
For a source ratio of 20
Transmissibility Ratio TR

0.5

0.4
Target Min. = 5
0.3
gives TR = .20
0.2

0.1

0
1 2 3 4 5 6 7 8 9 10

Stiffness Ratio; K body / K iso
Slide 57



Effective Isolation


Slide 58

Identifying Key NVH Paths
Key NVH paths are identified by Transfer Path Analysis (TPA)

Acoustic
Acoustic
Transfer
Transfer

Fi Break the system at the
points where the forces
enter the body (Receiver)

Operating loads Operating loads

Total Acoustic Response is summation of partial responses
over all noise paths

Pt = Σ paths [Pi ] = Σ paths [ (P/F) i * Fi ]
Slide 59

Designing Noise Paths
P/V Acoustic Transfer (P/F)i
Acoustic Transfer (P/F)i

P/F
V/F
F F (Kbody) F F
F F Fi F
Operating loads create
Forces (Fi) into body at
All noise paths
Pt = Σ paths [Pi ] = Σ paths [ Fi * (P/F) i ]
= Σ paths [ Fi * (P/V) i * (V/F) i ]
Measurement Parameters Generic Targets
P/F Acoustic Sensitivity 50 - 60 dBL/N

V/F Structural Point Mobility 0.2 to 0.3
(Receiver Side) mm/sec/N Slide 60

“Downstream” Effects: Body Panels
Recall for Acoustic Response Pt
Pt = Σ paths [Pi ] = Σ paths [ Fi * (P/V) i * (V/F) i ]
(P/V)i ! “Downstream” Dynamics (Body Panel) : Three Typical Effects:

1. Panel Stiffness
Increased Stiffness

3. Panel Acoustic Contribution

2. Panel Damping
Increased Damping

Slide 61

Generic Noise Path Targets
K body K source
> 5.0 > 20.0
K iso K iso

Structural
Mobility < 0.2 to 0.3 mm/sec/N

Acoustic 50 - 60
<
Sensitivity dBL/N

Primary: Minimize the Source Force
< 1.0 N

Panel Damping Loss Factor
> .10
Slide 62

Final Remarks on Mid Frequency Analysis

• Effective isolation at dominant noise paths is critical

• Reduced mobilities at body & source and softened
bushing are key for effective isolation
• Mode Separation remains a valid strategy as modes
in the source structure start to participate

• Other means of dealing with high levels of response
(Tuned dampers, damping treatments, isolator
placement at nodal locations) are also effective

Slide 63

• Introduction
• Case Studies Greg Goetchius
• Closing Remarks
Slide 64

Tool Box Demo Test Results
Toolbox Demo Noise Test Results
90
85

80

70 68
SPL (dBA)

63
61
60
55

50 47

40 39

30
1) Baseline: 2) Imbalance + 3) No Imbalance, 4) No Imbalance, 5) No Imbalance + 6) #5 + Absorption 7) #6 + Insulator
Imbalance, No Isolation No Isolation No Isolation + Isolation + Mat
Isolation Damping Damping
Slide 65

• Introduction
• Case Studies Kiran Govindswamy
• Closing Remarks
Slide 66

“Time Domain” Transfer Path Analysis
VINS (Vehicle Interior Noise Simulation)
Excitation Transfer Behaviour Interior Noise
Airborne
Loudspeaker Noise Share
Airborne Path

Microphones

FEV FEV
Overall
Power Train Noise Interior Noise
Intake&Exhaust Noise
...

Engine Mount Vehicle
Structure-Borne Path

Pi

Power Train Vibration
Drive Line Vibration Structureborne
Overall Vehicle Transfer Function
... Noise Share

Slide 67

VINS (Vehicle Interior Noise Simulation)
Overall Structure-Borne Sound Transmission

Resultant Dynamic Mount Stiffness

ae ab Fb

Mount Apparent Mass Vehicle Body
Engine Excitation Transmissibility of Body
Test Bench Impact Test

abody Fbody Pinterior
aengine = Pinterior
aengine abody Fbody
Resultant Dynamic Mount
Stiffness

Slide 68

VINS: Correlation (Full Load Speed Sweep)

Slide 69

VINS: Application to “Idle Modulation Noise”

Slide 70

VINS: Application to “Idle Modulation Noise”

Slide 71

VINS: Application to “Transient Diesel Clatter”
Measurement VINS VINS-SBN VINS-ABN

Slide 72

VINS: Application to “Transient Diesel Clatter”

Mic. left Mic. right Mic. top Mic. bottom Mic. front Mic. rear Intake orif. Exhaust tp.

Left mnt. Right mnt. RR A RR B Trans.
link

Slide 73

HEV, PHEV, ReEV, and EV NVH
Coasting G D C
Constant D C D C
G
speeds
A D
Acceleration B F
A D D
Drive-away C
B F
Start & A C F A F D
idle B E B G
Red light
E A
(Stop)
ICE ICE Start/Stop Regeneration Power AER
stopped idling ICE Batt. Charging Assist

• A: Global powertrain vibration B: Driveline vibration
• C: Noise of HEV specific components D: Magnetic noise E-Motor/(Generator)
• E: Noise of accessories F: Gearbox rattle noise due to
• G: Noise pattern changeover ICE (non-uniformity) and E-Motor

Slide 74

• Introduction
• Case Studies
Jianmin Guan
• Closing Remarks
Slide 75

Quietness during Idle and Electric-Vehicle Operation

Root Cause Diagnostics:
1. Water pump in the inverter cooling system
2. Electromagnetic noise of the motor,
the inverter, and other units

Reduce Inverter Water Pump Loads:
1. Reduce pump impeller imbalance
2. Redesign bearing structure
3. Changing motor structure

Improve Isolation from Body:
1. Install rubber isolator
2. Increase mounting bracket rigidity
3. Improve Inverter case

Slide 77

Quietness during Idle and Electric-Vehicle Operation

1. Water pump in the inverter cooling system
2. Electromagnetic noise of the motor,
the inverter, and other units

Reduce Inverter Water Pump Loads:
1. Reduce pump impeller imbalance
2. Redesign bearing structure Reduce
3. Changing motor structure Source

Effective
Isolation
Improve Isolation from Body:
1. Install rubber isolator Attach.
Stiffness
2. Increase mounting bracket rigidity
3. Improve Inverter case
Downstream
Slide 78

Engine Start Vibration

1. Engine torque fluctuations
2. Engine torque reaction forces

Reduce Torque Fluctuation:
1. Change intake valve closing timing
2. Control piston stop position
3. Adjust injected fuel volume and
ignition timing

Reduce Effect of Engine Loads:
1. Operating MG1 at high torque during engine start
2. Implement vibration-reducing motor control
3. Use two stage hysteretic torsional damper
4. Shorten distance between principal elastic
axis and center of gravity of power plant
Slide 79

Engine Start Vibration

1. Engine torque fluctuations
2. Engine torque reaction forces

Reduce Torque Fluctuation:
1. Change intake valve closing timing
2. Control piston stop position
3. Adjust injected fuel volume and
ignition timing
Reduce
Source
Mode
Manage.
Reduce Effect of Engine Loads:
1. Operating MG1 at high torque during engine start Damper
2. Implement vibration-reducing motor control
3. Use two stage hysteretic torsional damper
4. Shorten distance between principal elastic
axis and center of gravity of power plant Effective
Isolation
Slide 80

2nd Order Engine Induced Boom

1. 2nd order couple of the reciprocating inertia of piston
2. THS II Trans 50 mm longer and 35 kg heavier
3. Lower power plant bending mode
4. Requires 1.5X higher mount rates

Reduce Effect of 2nd order Couple:
1. Increase power plant bending mode
2. Move mount to a nodal point
3. Embed mounts inside cross member
4. Reduces distance from principle
elastic axis to CG
5. Optimized vertical to lateral rate ratio
Slide 81

2nd Order Engine Induced Boom

1. 2nd order couple of the reciprocating inertia of piston
2. THS II Trans 50 mm longer and 35 kg heavier
3. Lower power plant bending mode
4. Requires 1.5X higher mount rates

Mode
Manage.
Reduce Effect of 2nd order Couple:
1. Increase power plant bending mode Nodal
2. Move mount to a nodal point Mounting
3. Embed mounts inside cross member
Effective
4. Reduces distance from principle Isolation
elastic axis to CG
5. Optimized vertical to lateral rate ratio
Slide 82

Engine Radiated Noise

1. 24th MG2 order excitation
2. Lower MG2 reduction gear ratio
3. Lower transmission bending mode
- two key modes identified

Reduce 24th order Loads:
Arranged permanent magnets in V shape with optimized
angle

Reduce Effect of 24th order Loads:
1. Modified Trans case to improve modes
2. Added dynamic damper at a high amp. point
on Trans case
3. Added ribs in high radiating area of Trans case

Slide 83

Engine Radiated Noise

1. 24th MG2 order excitation
2. Lower MG2 reduction gear ratio
3. Lower transmission bending mode
- two key modes identified

Reduce
Reduce 24th order Loads: Source

Arranged permanent magnets in V shape with optimized
angle

Mode
Reduce Effect of 24th order Loads: Manage.

1. Modified Trans case to improve modes Damper
2. Added dynamic damper at a high amp. point
on Trans case
3. Added ribs in high radiating area of Trans case

Downstream
Slide 84

• Introduction
• Case Studies
• Closing Remarks Jianmin Guan Slide 85

Structure Borne NVH: Concepts Summary
• Source-Path-Receiver as a system
1.
1. Reduce Source
Reduce Source
2.
2. Rank and Manage Paths
Rank and Manage Paths
3.
3. Consider Subjective Response
Consider Subjective Response

• Effective Isolation
• Mode Management
• Nodal Point Placement
• Attachment Stiffness
• “Downstream” (Body Panel) Considerations

Slide 86

SAE 2011 NVH Conference

Thank You for Your Time!

Q&A

Slide 87

Structure Borne NVH References
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 Workshop - on Internet
At SAE www.sae.org/events/nvc/specialevents.htm
Slide 88
WS + Refs. at www.AutoAnalytics.com/papers.html

Structure Borne NVH References
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).

Slide 89

2011 SAE Structure Borne NVH Workshop Web

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