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1
SIMULATION OF VIBROACOUSTIC PROBLEM USING COUPLED FE / FE FORMULATION AND
MODAL ANALYSIS
Ahlem ALIA
presented by Nicolas AQUELET
Laboratoire de Mécanique de Lille
Université des Sciences et Technologies de Lille
Avenue Paul Langevin, Cité Scientifique
59655 Villeneuve d’Ascq, France
2
Introduction
Structure Fluid
The vibrations generate sound
The sound engenders vibrations
Main industrial concern in Vibroacoustics: Reduction of NOISE
Actually, noise constitutes an important indicator of quality in many industrial products such as vehicles, machinery…
3
Introduction
Analytical technique
Simple geometry
Under very restrictive hypothesis
Numerical methods
FEM / FEM
FEM / BEM
4
IntroductionClassical FEM / FEM
Six Nodes / Wavelength Application domain: Low
Frequency Range
DOF f
5
FEM / FEM with Modal analysis
Modal analysis solves the vibroacoustic problem with some modes
Reduction of the problem size
The modal analysis is applied with a Lumped mass representation
Lumped Mass matrix consists of Zero-off diagonal terms
Advantage of this approach:
Reduction of the computational cost
Introduction
(100 modes in our problem versus 432 physical unknowns)
6
Introduction Model the vibroacoustic behavior of an acoustic cavity with one flexible wall boundary by using FEM/FEM with:
Modal analysis
Lumped mass representation
Simply supported elastic plate
Mechanical load
Rigid wall
7
Governing equations Structure
Fluid
i
2
j
ij wx
0pkp 22 Fluid (f)
(f)
(sf)
Structure(s)
Vibroacoustic problem
0pnn ijij
0wnp
n
2
Pressure Continuity
Normal Displacement Continuity
BC at Coupling InterfaceP: pressure
k=/c wave number
w: displacement
: stress
n: interface normal
(1)
(2)
(3)
(4)
8
The application of the FEM to the variational formulation of structure cavity system yields to the following linear system :
Coupling system
Ks, Ms: structural stiffness and mass matrices
Kf, Mf: fluid matrices
B: coupling matrix
Fs, Ff: mechanical load, acoustical sources
c: sound velocity, f: fluid density
f
s
fT
s
f
s
F
F
p
w
MBc
M
K
BK2
20
0
MK F
p
wMK
2
,dnNNB
,dNNM,dNNM
,dNNcK,dNDNK
sfsTf
ffTf
ffssTs
ss
ffTf2
fssTs
s
Ns, Nf : structural and fluid shape function
(5)
(6)
9
The problem (6) can be seen as an eigenvalue problem:
(K - 2 M ) = F
p
w
Coupling system
For a great number of DOF, solving the system directly is always hard in term of CPU time.
(6)
10
Purpose of the approach
LFIRp
w 12
LKRILMR We search two matrices L and R verifying:
- L and R contain the LEFT and RIGHT eigenvectors, respectively
MRKR - is a diagonal matrix containing the eigenvalues of:
We obtain the physical unknowns of
MLKL HH
(K - 2 M ) = F
p
w
by this relation:
(7)
(10)
(8)
(9)
(11)(6)
11
Efficient eigenvalue algorithms can’t be used
A symmetric form of eigenvalue problem is required
(K - 2 M ) = F
p
w
Since
is non-symmetric,
Sandberg’s method enables us to make it symmetric by using Modal analysis
Purpose of the approach
(6)
12
Classical modal analysis
fffff XMXK sssss XMXK
s , f : the structural and the fluid eigenvalues.
Xs , Xf : the structural and the fluid eigenvectors.
Cavity with stiff boundaries Structure in vacuum
fXsX
Solved
Independently
(12) (13)
13
Classical modal analysis
f
s
f
s
p
w
0
0
fffTffff
Tf
sssTssss
Ts
DK,IM
DK,IM
Ds , Df are diagonal matrices containing the structural and the fluid eigenvalues.s, f represent the modal structural displacement and the modal fluid pressure.
fXsX
s, f are matrices containing some eigenvectors of structure and fluid
s, f verify the following properties:
(14)
(15)
(50 modes) (50 modes)
14
Classical modal analysisHence, the coupling system (5) can be rewritten as the reduced system (17):
f
s
f
s
p
w
0
0 s, fw , p
f
s
fT
s
f
s
F
F
p
w
MBc
M
K
BK2
20
0
fff
sTs
f
s
f2
fsTT
f22
fTss
2s
F
F
IDBc
BID
The system is reduced (the problem size is divided by 4) but it remains non symmetric
(5)
(14)
(17)
432 physical unknowns 100 modal unknowns
15
Modal analysis (Sandberg Method)
fff
sTs
f
s
f2
fsTT
f22
fTss
2s
F
F
IDBc
BID
f
s
f
s
f
s
f
sS
I
Dc
0
012
sTss
TTff
Tf
sTss
f
s
fTss
TTffss
TTf
fTsss
FBcF
FDc
BBcDDBc
BDcDI
2
2
22
22
Symmetric system
Non Symmetric system
Transition matrix
(17)
(18)
(19)
16
sTss
TTff
Tf
sTss
f
s
fTss
TTffss
TTf
fTsss
FBcF
FDc
BBcDDBc
BDcDI
2
2
22
22
Symmetric generalized eigenvalue problem
FAIF
S
2
VAV
Modal analysis (Sandberg Method)
V : right eigenvector matrix of the symmetric system
(19)
(20)
(21)
17
V
I
DcRf
s
f
s
0
00
01
2
f
s
f
s
p
w
0
0s, f
w , p
f
s
f
sS s, f
Modal analysis (Sandberg Method)
R: right eigenvector matrix of the original system
The left eigenvector matrix “ L” is obtained in the same manner
MRKR
(18)(14)
(22)
(9)
18
Modal analysis (Sandberg Method)
(K - 2 M ) = F
p
w
LFIRp
w 12
LKR
ILMR &
is a diagonal matrix containing the eigenvalues of:
(6)
(11)
MRKR MLKL TT (10)
(9)
19
Numerical results
Coupling
InterfaceFluidRigid
Structure
Elastic Structure
20
Numerical results
Cavity
(c)
Plate
(b)
Discrete Kirchhoff Quadrilateral (DKQ) plate element thin plate
Kirchhoff theory
8-node brick isoparametric acoustic element
21
Structure ( Frequency response) Simply supported plate (0.5m 0.5m)
Unit punctual force (0.125m 0.125m)
Variation of the displacement with the frequency at the load point
Results given by Migeot et al (1)Numerical results
(1) 2nd Worldwide Automotive Conference Papers,1-7
22
Structure ( Natural frequencies)Structure: Simply supported plate (0.2m 0.2m) made of brass
0 20 40 60 80 1000
1000
2000
3000
4000
5000
6000
Mode number
Fre
que
ncy
(Hz)
Analytical Consistent mass Lumped mass
Natural frequencies of the plate
Consistent and lumped mass matrices are in good agreement with analytical ones as long as low frequencies are considered (<50th mode).
23
Cavity ( Natural frequencies)
0 20 40 60 80 1000
2500
5000
Mode number
Fre
que
ncy
(Hz)
Analytical FEM
Rigid cavity (0.2m 0.2m 0.2m)
FEM leads to good results below the 50th mode
Natural frequencies of the rigid cavity
24
Coupling problem
1 0 0 0- 8 0
- 6 0
- 4 0
- 2 0
0
2 0
F r e q u e n c y ( H z )
Pr
es
su
re
( d
B )
D i r e c t F E M M o d a l F E M w i t h c o n s i s t e n t m a s s M o d a l F E M w i t h l u m p e d m a s s
10 100 1000-100
-80
-60
-40
-20
0
20
40
Frequency ( Hz )
Pre
ssure
(dB
)
Direct FEM Modal FEM with consistent mass Modal FEM with lumped mass
Mesh (1266)Direct Method Modal analysis32mn 54s 56s
Results Given by Lee et al (2)
CPU Time
Simply supported elastic plate
Field point
Pressure at the point (0.1,0.1,0.2)
Numerical results
(2) Engineering Analysis with Boundary Elements, 16 (1995) 305-315
25
Structure ( Frequency response)
200 400 600 800 1000-200
-150
-100
-50
0
50
Frequency (HZ)
Dis
plac
emen
t (dB
)200 400 600 800 1000
-150
-100
-50
0
50
Frequency ( Hz )
Dis
plac
emen
t (dB
)
Plate quadratic displacement of the structure
In vaccum Plate-cavity (air)
Coupling effect
854Hz
26
Mea
n sq
uare
pre
ssur
e
Frequency (Hz)
Mea
n sq
uare
vel
ocity
Frequency (Hz)
Coupling problem (air)
Comparison between the direct and the modal results
Mean square pressure: cavity Mean square velocity:structure
27
Coupling problem (water)
Comparison between the direct and the modal results
Mean square pressure: cavity Mean square velocity: structure
Mea
n sq
uare
vel
ocity
Frequency (Hz)
Mea
n sq
uare
pre
ssur
e
Frequency (Hz)
Mea
n sq
uare
vel
ocity
28
FEM / FEM -- FEM / BEM
FEM-FEM FEM-BEM comparison
Frequency (Hz)
Pre
ssu
re (
dB
)
Simply supported elastic plate
Field point
29
Conclusion FEM / FEM with modal analysis and lumped mass representation has been used to model a simple vibroacoustic problem.
A good representation of the mass is very essential to achieve accurate results.
Modal FEM / FEM with only small number of modes is less efficient for strong coupling.
More modes must be taken into account ( disadvantage)
Solution: Improve the numerical results by using Modal correction for diagonal system