For Peer Review

Interior structure-borne noise reduction by controlling

automotive body panel vibration

Journal: Journal of Vibration and Control

Manuscript ID: JVC-10-0339

Manuscript Type: Original Manuscript

Date Submitted by the Author:

21-Dec-2010

Complete List of Authors: guo, rong zhang, lijun zhao, jing zhou, hong

Keywords: Structure-borne noise, FEM-BEM, TPA, PACA, Damping treatment

Abstract:

The adoption of numerical method to perform the vibration and acoustic analysis is attracting increasing attention because of its merits of saving costs and time. The analysis flow chart of structure-borne noise reduction by controlling vehicle body panel vibration is brought forward. According to this flow chart, finite element analysis of body in white is performed by MSC. Nastran and the experimental modal analysis of body in white is also carried out. To verify finite element model, the correlation analysis between test and FE results is done. After the acoustic boundary element model with seats is built, the acoustic transfer vector (ATV) is calculated by using LMS software. The transfer path analysis is performed to estimate the powertrain and exhaust pipe operational

forces at idle. Then the forces are inputted into the acoustic BEM to calculate body panel velocity. By combining ATV and body panel velocity, the interior acoustic response can be obtained. The simulation results are verified by the test. To point out the main contribution panels for the interior sound pressure level (SPL), panel acoustic contribution analysis (PACA) is based on BEM. At last, constrained damping treatments are adopted to control the main panel vibration based on PACA results. Test results indicate that the A-weighted SPL of interior noise at idle is reduced obviously before and after treatments, which has proved the correctness of PACA results. So it is believed that the proposed analysis method is feasible for reducing interior structure-borne

noise.

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Interior structure-borne noise reduction by controlling automotive

body panel vibration

Rong Guo a,*

, Li-jun Zhang a, Jing Zhao

b, Hong Zhou

a

(a

Tongji University, National engineering laboratory of new energy vehicles and power systems, Shanghai, China)

(b

Shanghai Volkswagen Automotive Company Ltd., Shanghai, China)

Abstract

The adoption of numerical method to perform the vibration and acoustic analysis is attracting increasing

attention because of its merits of saving costs and time. The analysis flow chart of structure-borne noise

reduction by controlling vehicle body panel vibration is brought forward. According to this flow chart, finite

element analysis of body in white is performed by MSC. Nastran and the experimental modal analysis of

body in white is also carried out. To verify finite element model, the correlation analysis between test and FE

results is done. After the acoustic boundary element model with seats is built, the acoustic transfer vector

(ATV) is calculated by using LMS software. The transfer path analysis is performed to estimate the

powertrain and exhaust pipe operational forces at idle. Then the forces are inputted into the acoustic BEM to

calculate body panel velocity. By combining ATV and body panel velocity, the interior acoustic response can

be obtained. The simulation results are verified by the test. To point out the main contribution panels for the

interior sound pressure level (SPL), panel acoustic contribution analysis (PACA) is based on BEM. At last,

constrained damping treatments are adopted to control the main panel vibration based on PACA results. Test

results indicate that the A-weighted SPL of interior noise at idle is reduced obviously before and after

treatments, which has proved the correctness of PACA results. So it is believed that the proposed analysis

method is feasible for reducing interior structure-borne noise.

Key words: Structure-borne noise; FEM-BEM; TPA; PACA; Damping treatment

1. Introduction

Nowadays, one of the most valuable criteria for vehicle quality assessment is based on acoustic

emission levels: the comfort of a car depends on the noise level transmitted inside. Consequently, there is a

general attention to design criteria aiming at improving the structural-acoustic behaviour, in such a way to

withstand the increasingly restrictive ergonomic standard [1]. Interior noise can be divided into airborne

noise and structure-borne noise according to their transmission modes. Usually the noise below 500Hz is

transmitted by structure or solid. The structure-borne noise can be reduced by controlling noise sources or

their transfer paths.

For structure-borne analysis the frequency range of interest below 250Hz is determined by the major

sources of input energy. The excitation due to road conditions is typically predominant at frequencies less

than 25Hz, and it is usually random in nature and decrease rapidly with increasing frequency. The powertrain

idle and its harmonics have a strong excitation of frequencies below 250Hz [2]. This objective of this study

is to reduce interior structure-borne noise by controlling the body panel vibration without adjusting

powertrain mount system and suspension system.

* Corresponding author. Tel: +86 13564030444 Fax: +86 021-69589843

E-mail addresses: guorong@tongji.edu.cn

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The structure-borne noise can be studied by numerical methodologies such as boundary element method

(BEM) and finite element method (FEM) that are effective up to medium frequencies [1]. The combined

usage of BEM and FEM is becoming more and more popular for vibro-acoustic analysis, with FEM to be

preferred for the structural dynamic simulation and BEM for the acoustic field assessment [1-8]. Such

combined approach provides results that are useful to plan active and passive design adjustments to minimize

the structure-borne noise that is transmitted inside the car body, so as to avoid very costly experimental

campaigns.

The analysis method and flow chart of structure-borne noise reduction by controlling body panel

vibration is brought forward in the present study. The finite element method is used to calculate the structural

characteristics of body in white (BIW) and correlation between FE and test results are done. After the

acoustic boundary element model with seats is built, the acoustic transfer vector (ATV) is calculated.

The transfer path analysis is performed to estimate the powertrain and exhaust pipe operational forces at

idle. Then the forces are input into the acoustic BEM to calculate body panel velocity. By combining ATV

and body panel velocity, the interior acoustic response can be achieved. And the simulation results are

verified by the test. To point out the main contribution panels for the interior SPL, panel acoustic

contribution analysis (PACA) is done based on BEM. At last, constrained damping treatments are adopted to

control panel vibration based on PACA results, and the test results verify the correctness of PACA.

2. Analysis method of structure-borne noise reduction by controlling body panel

vibration

2.1 Analysis flow chart of structure-borne noise

Fig. 1 Analysis flow chart of structure-borne noise reduction by controlling body panel vibration

To predict and control interior noise, the analysis flow chart of structure-borne noise reduction based on

FEM-BEM method is proposed as is shown in Fig. 1. First, the finite element of body in white (BIW) is

modeled. To determine the whole modes and important local modes, the finite element analysis of BIW is

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performed, and the experimental modal analysis of body in white is also carried out. The correlation analysis

between FE and test results is performed to verify finite element model. Then, the excitation forces obtained

from transfer path analysis (TPA) are loaded at the specific locations (powertrain mounts and exhaust pipe

hangers) of BIW to get body velocity response as the boundary conditions of panel acoustic contribution

analysis. On the other hand, the boundary element model of body with seats is modeled. After setting up field

points in BEM, the acoustic transfer vectors can be calculated. Then, to point out the direction for body panel

improvement, the panel acoustic contribution analysis (PACA) of peak frequencies at interior locations is

carried out through combining body velocity responses and ATVs. At last, constrained damping treatments

are adopted to control body panel vibration based on PACA results and the interior noise improvement is

verified by the test results.

2.2 Theory

The adopted FEM-BEM approach takes advantage of the Acoustic Transfer Vector (ATV) algorithm that

is particularly useful when big problems are to be solved [9]. ATV provides input-output relations between

the node normal velocities , ( )e iv ω of the radiating surface and the sound pressure level ( )p ω at a specific

field point (Eq. 1).

,

1

( ) ( ) ( )N

i e i

i

p ATV vω ω ω=

= •∑ （1）

where the terms N and ω respectively represent the total number of elements and angular frequency.

, ( )e iv ω is the normal velocity of the ith element.

ATVs can be interpreted as an ensemble of acoustic transfer functions from the surface nodes to a single

field point and they only depend on the configuration of the acoustic domain, i.e. geometry and properties

(speed of sound and density), on the acoustic surface treatment, on the frequency, and on the field point; they

do not depend on the loading conditions and structural response.

Body panel vibration velocity varies with different driving conditions. The acoustic response is

calculated for all loading conditions by combining the ATVs with the normal structural velocity boundary

condition vectors. However, acoustic transfer function is evaluated from the radiating surface to the specific

field point for the involved range frequency, without taking into account the actual loading conditions. When

the acoustic domain has not been changed, to shorten the time for solving the acoustic response and to

improve the analysis efficiency, we could use ATV technology to predict acoustic response in different

driving conditions.

3 Structural finite element analysis and test verification

3.1 Finite element analysis of BIW

Hypermesh software is used to mesh BIW with triangle and quadrilateral shell elements as shown in Fig.

2. The pre-treatments in Hypermesh include: (1) accurately simulate the connection features between the

sheet metal and other connections and location of the solder joint. The RBE2 element is used to simulate

weld spot in the present study. (2) accurately reflect the thickness of the sheet metal parts, material properties

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and other information.

Fig. 2 Finite element model of body in white

After the finite element has been modeled, the MSC/NASTRAN SOL103 is used to calculate the

structural modes of BIW [5].

3.2 Experimental modal analysis of BIW

3.2.1 Test system

The test system consists of three parts: exciter system, data acquisition system and modal analysis and

process system. Exciter system includes QDAC signal generator module of LMS SCADASⅢ SC316W,

power amplifiers and shakers. Data acquisition system includes acceleration transducers, force transducers

and LMS SCADASⅢ SC316W signal amplification and acquisition system. The software used in modal

analysis is LMS Test.lab. The illustration of the whole system is shown in Fig. 3.

Fig. 3 Schematic diagram of the modal test system

Fig. 4 Geometry of BIW in LMS Test.lab Fig. 5 Suspension method of BIW

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3.2.2 Test method

The “Multiple Input-Multiple Output” method is adopted in the modal test. 145 measuring points in

three directions are measured in the modal test. The geometry of BIW in LMS test.lab is shown in Fig. 4.

Finite element calculations of BIW show that the horizontal (y) and vertical (z) modes are the main modes,

while the longitudinal (x) modes in the low frequency range are the minor. Therefore, in order to ensure the

horizontal and vertical modes adequately excited, two points in z direction and one point in Y direction are

selected as driving points as shown in Fig. 4. From Fig. 5 it is seen that the tested BIW is suspended by

elastic ropes to ensure suspension frequency below the 10 percent of the first mode of BIW.

3.3 Correlation analysis between test model and finite element model

Correlation analysis can help to quantify the geometrical and dynamic (FRF and Modal) resemblance

between the test model and its FE equivalent model. For modal analysis, it means the degree of correlation

between modal parameters of FE and test. In general the test model is always considered more reliable. So

the correlation analysis is often used to evaluate and to improve the accuracy of theoretical models. The

related criteria of modal analysis mainly include frequency criterion, Modal Assurance Criterion (MAC),

Coordinate Modal Assurance Criterion (COMAC), Frequency Response Assurance Criterion (FRAC), and

Response Vector Assurance Criterion (RVAC) [10]. For modal analysis, the first concern is the modal

frequencies and modal shapes.

3.3.1 Frequency criterion

In general the precision of measured frequency t

ω is the highest in modal parameter estimation. The

difference between t

ω and calculated frequency c

ω denotes degree of correlation between them, usually

expressed as a percentage:

t c

t

ω ωω−

（%） （2）

Table 1 Comparison of natural frequencies obtained from FE and test

Test (Hz) FE (Hz) Relative error (%) Modal shape description

24.6 23.6 4.1 global shearing movement in Y direction

27.2 29.6 -8.8 roof local mode

33.0 32.5 1.5 1st global torsional mode

37.0 37.9 -2.4 roof local mode

37.9 36.4 3.9 front crossbeam local mode

40.3 40.3 0 twist around z axis and roof local mode

40.9 40.9 0 global surge

43.5 43.7 -4.5 global surge and roof local mode

43.8 43.4 0.9 global shearing movement in X direction

48.7 48.3 0.8 global surge and roof local mode

58.9 59.2 -0.5 1st global bending mode and roof local mode

Table 1 is the comparison of natural frequencies obtained from FE and modal test. It is seen from the

table that most of the errors are within 5%, in particular, the errors of main global modes are small.

Furthermore, the vibration modes of FE and modal test are also consistent.

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The comparison of modal shapes obtained from FE and test is shown in Fig. 6. The left and right part

respectively denote the FE results and test results. From Fig. 6 it is seen that the modal shapes of FE and test

agree well.

(a) 23.6Hz

(b) 32.5Hz

(c) 40.3Hz

(d) 40.9Hz

(e) 43.4Hz

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（f）48.3Hz

Fig. 6 Comparison of modal shapes obtained from FE and test

3.3.2 Modal assurance criterion

Although the consistency of mode shapes can be evaluated by observing the modal shapes of FE and

test, this comparison is only qualitative analysis. Through the calculation of MAC, the degree of relationship

between modal shapes of FE and test can be quantitatively determined, and thus a more objective evaluation

can be given. The MAC values always lie between 0 and 1. The value of ijMAC is 0 when the FE and test

model are orthogonal, and 1 when the two models are equal. So, the corresponding test and FE modes should

have a high MAC values. The MAC matrix is built on position ij, using the test modal vectors ( }{ test

iψ ) and

FE modal vectors ( }{ FE

jψ ):

} }{{}{ }{( }{ }{(

2*

* *

test FE

i j

ijtest test FE FE

i i j j

MAC

ψ ψ

ψ ψ ψ ψ=

（3）

By the formula 3 we can see that the calculation of MAC should only consider the global mode shape

information. If the local modes are included in above mode shapes, the results will affect the credibility of

MAC values. Therefore, when calculating the MAC values of BIW, we should filter out the global modes as

calculation objects.

Table 2 MAC values of the main modal shapes of FE and test

Test FE 23.6Hz 32.5 Hz 40.3 Hz 40.9 Hz 43.4 Hz 48.3 Hz

24.6 Hz 0.93 0.02 0.02 0.01 0.00 0.02

33.0 Hz 0.01 0.93 0.00 0.01 0.02 0.01

40.3 Hz 0.01 0.00 0.79 0.15 0.15 0.10

40.9 Hz 0.04 0.02 0.12 0.71 0.01 0.00

43.8 Hz 0.02 0.01 0.01 0.02 0.57 0.09

48.7 Hz 0.07 0.00 0.00 0.03 0.00 0.75

Table 2 shows MAC values of the main few global modes. From the table, it shows that the values of

non-diagonal elements are close to 0 and those of the diagonal elements are above 0.7. This indicates that the

modal shapes of FE and test agree well. Thus the FE model has a high accuracy calculation for the next

analysis.

For the vibration mode of 43.8 Hz, the MAC value is small and does not meet the basic requirements

(above 0.7). This is due to the modal shapes of this frequency for the dislocation in x direction. At the same

time, in order to ensure the horizontal and vertical modes adequately excited in modal test, there are no

driving points in X direction. Thus this type of excitation could not provide enough excitation energy in X

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direction which causes the error.

4 Acoustic boundary element modeling

Because the existence of the seat has a great influence on the results of boundary element model [6], the

seats are included in the BEM. First, the CAD model of three rows of seats is modeled by using CATIA

software. Then the seat models are meshed in Hypermesh as shown in Fig. 7.

Through the LMS's automatic mesh generator, the acoustic finite element model of 80mm mesh size is

developed. According to the principle that there are at least six gird elements in a wavelength [7], the upper

analyzed frequency will reach to the 708.3Hz. The frequency can meet the research requirements of the

present paper. The acoustic boundary element model with seats is shown in right part of Fig. 7.

Fig. 7 Acoustic boundary element model

5 Interior structure-borne noise analysis

To determine the acoustical response, one must first compute acoustic transfer vector by the boundary

element (BE) method and determine the velocity response for the frequency range of 20-200 Hz on the

structure of the cab by the finite element (FE) method. The excitations for the velocity response using FEM

are the forces at the cab side of the mounts such as powertrain mounts and exhaust pipe hangers. This

velocity response is then translated into a boundary condition for the prediction of interior SPL by BEM.

5.1 ATV calculation

Before ATV calculation, the location of driver right ear (DRE) is set to the field point. And the property

of material wrapping the body structure is set to air, while the seat material is usually foam.

The frequency range of interest in ATV calculation is 20-200Hz and the frequency step is 1 Hz. The

ATV calculation results at DRE are shown in Fig. 8. ATV results are expressed as contours, and the different

colors of element correspond to the ATV values.

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(a) 27Hz (b) 54Hz

(c) 118Hz (d) 135Hz

Fig. 8 ATV calculation results at DRE

5.2 Structural frequency response analysis

5.2.1 Operational force calculation at idle

Engine idle condition is studied in the present paper. The engine mount exciting force transfer of XYZ

three translational degree of freedoms (DOFs) is considered and the moment transfer of rotational DOFs are

ignored. So there are nine transfer paths for three points engine mount. And we assume that the force

transmission of these paths is the process of energy coupling to each other. For exhaust pipe hangers, because

the exciting forces in X and Y direction have less influence on the interior noise, the forces in Z direction are

considered. Similar to engine mounts, we assume that the force transmission of four hangers is also the

process of energy coupling to each other. The matrix inversion method [11] is used to identify the exciting

forces. Test plans of transfer path analysis (TPA) are as follows:

(1) Two triaxial accelerators should be installed in body side of each powertrain mount. Hammer is used

to impact body side (as close to connection points as possible) of each mount in XYZ directions. At the same

time the force signal and the acceleration signal should be recorded.

(2) Two one-way accelerators should be installed in body side of each exhaust pipe hanger. Hammer is

used to vertically impact body side (as close to connection points as possible) of each hanger in Z direction.

At the same time the force signal and the acceleration signal should be recorded.

(3) Open the engine and let the engine warm up for a period of twenty minutes. After the engine idle

condition is stable, each accelerator signal should be recorded.

(4) The microphone is installed at DRE to measure the acoustic pressure at idle.

According to the matrix inversion method and above test plans, the exciting force formulas of engine

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mounts and exhaust pipe hangers are shown in Eq. 4 and Eq. 5. To improve the accuracy of force

identification, nine reference DOFs in powertrain mounts and four reference DOFs in exhaust piper hangers

are designed in TPA.

[ ] [ ] [ ] 1

1 9 1 18 18 9PT PT PTF R H−

× × ×= × （4）

[ ] [ ] [ ] 1

1 4 1 8 8 4EP EP EPF R H−

× × ×= × （5）

where the terms [ ]iF and [ ]iR respectively represent exciting force vectors and response vectors.

The term [ ]iH is transfer function maxtrix. ‘i = PT’ and ‘i = EP’ respectively denote powertrain and

exhaust pipe.

(a) left hand side mount (b) right hand side mount

(c) transmission side mount (d) exhaust pipe mount

Fig. 9 Operational force calculation at idle

To avoid the errors enlarged after matrix inversion, singular value decomposition of the transfer

function matrix (Eq. 4 and 5) is done to exclude the small singular values. The exciting force estimation of

PT mount and EP hanger is shown in Fig. 9 based on Eq.4 and 5.

Fig. 9 shows that the exciting force of right hand side (RHS) mount is the biggest and the one of

transmission side mount is relatively smaller. Furthermore, the exiting forces of RHS mount in X and Z

direction are obviously lager than those in Y direction. For exhaust pipe hangers, the fourth hanger force is

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the biggest in all 4 hangers and the third hanger force is the second.

5.2.2 Body panel velocity under operational force input

The powertrain mount system in present study is standard drive type with the characteristics of a

longitudinal engine, three points support and rear axle drive. The primary paths of structural energy into the

cab at idle include three PT mounts and four EP hangers. These mounts and hangers are used to provide

vibration isolation for the cab by reducing the transmission of energy into the cab. To predict the

structure-borne noise caused by PT mounts and EP hangers, the operational forces at idle are applied to the

location of these mounts and hangers. The body panel velocity response is shown in Fig. 10.

(a)27Hz (b)54Hz

(c)118Hz (d)135Hz

Fig.10 Body panel velocity response

5.3 Interior acoustic prediction due to structural excitations

This body velocity response in section 5.2 is translated into a boundary element in section 5.1 for the

prediction of interior SPL at DRE. Comparisons of simulation and test results of interior SPL are shown in

Fig. 11. It can be seen from the figure the peak positions of simulation and test, especially the main peak

frequencies of sound pressure are more consistent with the trend. Therefore, the model has a relatively higher

accuracy for further analysis.

It is concluded that the amplitude errors in some frequencies are mainly from the airborne noise

included in test and not in simulation. Also sound-absorbing material is not considered in the model.

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Fig. 11 Acoustic prediction at DRE due to structural excitation at powertrain mounts and exhaust pipe hangers

The engine in present study is a four-stroke engine with speed at idle is 800 r/min. So the fundamental

frequency is 13.33Hz and the main excitation frequencies for this type of engine are the even orders. It can

be seen from the figure that the main peak frequencies of interior SPL are 27 Hz (2nd order), 54 Hz (4th

order), 80 Hz (6thorder), 118 Hz (no order) and 135Hz (10th order). In addition to 118Hz, the noise spectrum

reflects the significant features of harmonic.

6 Body panel acoustic contribution analysis

6.1 Acoustic contribution coefficient

The concept of acoustic contribution coefficient is introduced to quantify each panel acoustic

contribution. Vibration of the eth element contribution to acoustic contribution of a field point De is the

projection of sound pressure Pe in total sound pressure P. Its expression is [12]:

*

e

e e

P PD R

P= (6)

where *P is the complex conjugate of P and

eR is real part of the complex

*

eP P

P.

It is seen in Eq. 7 that the total sound pressure c

P caused by this panel vibration is sum of SPL e

P

generated by each element vibration.

{ }[ ]c e e e

pe pe

P P ATV v= =∑ ∑ (7)

where pe is the number of elements making up the panel.

From Eq. 7, we can get panel acoustic contribution as follows:

*

c

c e

P PD R

P= (8)

It is seen from the definition of acoustic contribution coefficient that there are positive, negative and

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neutral acoustic contribution. If the phase angle difference between a panel pressure and total pressure is less

than 90 °, the panel or element contributes to the total amount in a positive way. The total pressure increases

with the increasing contribution. If the phase angle difference between a panel pressure and total pressure is

more than 90 °, the panel or element contributes to the total amount in a negative way. Their contribution

actually reduces the total pressure. If the phase angle difference between a panel pressure and total pressure

is close to 90 ° or the panel pressure amplitude is small, the panel or element contribute to the total amount in

a neutral way. Neutral contribution panel can not be modified to achieve the desired objectives. So it is very

important to identify positive and negative acoustic contributions for noise reduction before structural

modifications.

6.2 Panel acoustic contribution analysis

In order to do panel acoustic contribution analysis (PACA) the BEM model of cab is split into 22 panels

as shown in Fig. 12. Correspondence between the numbers and panels is shown in Table 3.

Fig. 12 Panel divisions for panel acoustic contribution analysis

Table 3 Correspondence between the numbers and panels

No. panel No. panel No. panel No. panel

1 front roof 7 left_front_door 13 window of right_mid door 19 firewall

2 middle roof 8 left_mid_door 14 window of right_rear side wall 20 front floor

3 rear roof 9 left_rear side wall 15 right_front_door 21 middle floor

4 window of left_front door 10 rear window 16 right_mid_door 22 rear floor

5 window of left_mid door 11 rear door 17 right_rear side wall

6 window of left_rear_side_wall 12 window of right_front door 18 front windshield

The PACA is done based on the formula of section 6.1. The statistical results of PACA are shown in

Table 4. Blue box, space and red box respectively represent positive, neutral and negative acoustic

contribution. From table 4 it is seen that panel 3, 9, 17, 18, 19, 20 and 21 contribute to total pressure in a

positive way for some frequencies. Panel 8 have positive or negative in different frequencies. Panel 2, 6 and

16 contribute to 27 Hz at the DRE in a negative way. Panel 1, 4, 5, 7, 11, 12, 13, 15 and 22 are neutral

contribution panels.

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Table 4 Panel positive and negative contribution at resonance frequencies

6.3 Test verification

In order to verify the accuracy of PACA in section 6.2, panel 17 (right_rear side wall) vibration is

controlled since this panel contributes the sound pressure of 54 Hz and 135 Hz at idle in a positive way. And

the sound pressure levels before and after modifications are compared.

Improving structure stiffness and damping are the two ways to suppress the vibration of body panel.

Due to high damping characteristics of visco-elastic materials, damping treatment in the structure can lower

the resonant response in very broad frequency range and generally not have significant additional peaks.

Therefore the damping treatment is more suitable for verifying the PACA conclusions in this section.

Because the constrained damping treatment can dissipate more energy and be laid more conveniently

compared to the free damping treatment [8], it is used to lower the panel vibration. Its substrate materials and

constrained layers are the butyl rubber and aluminum sheet.

Based on the PACA results, the constrained damping treatment is imposed on panel 17. The right_rear

side wall acceleration after damping treatment is shown in Fig. 13. It is seen from the figure that the panel

vibration is reduced obviously, especially in the frequency range of 100-160 Hz.

Fig. 13 Right_rear side wall acceleration after constrained damping treatment

The SPL at DRE after constrained damping treatment is shown in Fig. 14. From the figure it is seen that

the interior SPL in certain frequencies reduce obviously, for example 1.7 dB(A) reduction in 54Hz and 7

dB(A) in 135Hz. This conclusion is consistent with the PACA conclusions that panel 17 contributes to above

frequencies in a positive way.

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Fig. 14 SPL at DRE after constrained damping treatment

7 Conclusions

In the present study the FEM-BEM method is adopted to study the mini-bus structure-borne noise. The

analysis method and flow chart of structure-borne noise reduction by controlling body panel vibration is

brought forward and verified by test. The research work and conclusions are as follows:

(1) The finite element of BIW is modeled and the body structure modes are calculated. The correlation

analysis between FE and test results shows that the frequency relative errors of FE and test are small and the

modal shapes of FE and test agree well, which has proved the reliability of the finite element model.

(2) The acoustic boundary element model with seats is built. The location of driver right ear (DRE) is

set to the field point and the ATV in the frequency range of 20-200Hz is calculated by using LMS software.

(3) The matrix inversion method is used to estimate PT and EP operational forces at idle. Then the

forces are input into the acoustic BEM to calculate body panel velocity. By combining ATV and body panel

velocity, the interior acoustic response can be obtained. Comparison between simulation and test results

indicates that the main peak frequencies are more consistent with the trend.

(4) The BE model of cab is split into 22 panels and PACA is done based on the formula. The PACA

results point out the main contribution panels for peak values of SPL. By imposing constrained damping

treatment on the right_rear side wall, the panel vibration and interior SPL reduce obviously, which further

verifies the correctness of PACA conclusions.

The analysis method of structure-borne noise reduction is studied and the test results show that the

interior SPLs reduce significantly. So this analysis method and flow chart are expected to help automotive

sound engineers to reduce interior structure-borne noise.

Acknowledgments

We extend our gratitude to HAIMA Automobile Co., LTD. for their support of the project “HAIMA Z1

mini-bus NVH optimization”, and we also thank the tongji university for their support of the project of

“Program for Young Excellent Talents in Tongji University”.

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