Application of Acoustic-solid Coupling Theory in New Energy Vehicle
Noise Control
FUJUN MAO
Lyceum of the Philippines University,
Manila 1002,
PHILIPPINES
Abstract: - The development of new energy vehicles has attracted much attention due to the strong promotion
and popularisation of the concept of low carbon and environmental protection, and the increasing demand for
environmental protection in cars. Although these vehicles meet people’s requirements for resource and
environmental protection, the noise generated during the driving process affects the comfort of the vehicle
occupants and the concentration of the vehicle driver. To address this problem, the research proposes to
improve the noise control technology of new energy vehicles based on acoustic-solid coupling theory and to
test the practical application effect of this technology. The test results show that the maximum acceleration of
vibration at the roof, floor, axle head, and spring of the new energy vehicle are 1.48 m/s2, 1.02 m/s2, 0.079
m/s2, and 0.020 m/s2 respectively, which are lower than the maximum acceleration before the use of this
technology. The maximum sound pressure at the windscreen and side window glass of the new energy vehicle
is 80 dB(A) and 73 dB(A) after the use of this technology. The maximum sound pressure at the driver’s ear was
62 dB(A) and 77 dB(A) when the vehicle was driven on different road surfaces, which were lower than the
sound pressure values before use. In summary, the research proposes to improve the noise control technology
of new energy vehicles based on the sound-solid coupling theory, which can have the effect of reducing the
noise value generated by new energy vehicles and improving the comfort of users.
Key-Words: - New energy vehicles; Acoustic-solid coupling; Automotive noise; Noise control technology
Received: March 21, 2023. Revised: August 22, 2023. Accepted: September 23, 2023. Published: October 26, 2023.
1 Introduction
With the continuous development of science and
technology, many new technologies are being used
in many fields, and the automotive industry is
developing very fast thanks to these technologies,
[1]. With the gradual strengthening of the
performance of cars, people’s requirements for cars
are no longer limited to their safety, [2], [3]. At the
same time, the growing popularity of environmental
protection theory has influenced people’s demand
for environmentally friendly cars. Based on this
development trend, the automotive industry has
developed new energy vehicles that do not use
traditional fossil energy and reduce the consumption
of non-renewable energy sources, which are very
popular among people, [4], [5]. However, these
vehicles generate a lot of noise during the driving
process, which not only affects the comfort of the
occupants and the driver’s concentration but also
produces noise pollution to the surrounding
environment. To address this problem, the study
proposes the use of acoustic-solid coupling theory to
optimize the noise control technology of new energy
vehicles. It is expected that this method will
effectively control the noise during the driving
process of new energy vehicles and reduce the
negative impact of noise. The innovation of this
study is to use the acoustic-solid coupling theory,
combined with the structural modal analysis of the
new energy vehicle, to construct an acoustic-solid
coupling system model to accurately monitor the
noise inside the vehicle. The contribution of this
study is that through in-depth research and
exploration of the application of sound-structure
coupling theory in the noise control of new energy
vehicles, it provides a new idea and method for
theoretical research in the field of noise control of
new energy vehicles. By studying the noise sources
of new energy vehicles, targeted noise control
measures are proposed to reduce the internal noise
of new energy vehicles, improve the riding comfort,
improve the riding environment, improve the
comfort and satisfaction of drivers and passengers,
promote the promotion and popularization of new
energy vehicles, and make contributions to
environmentally friendly transportation. The first
part of this study is a brief introduction to recent
research on automotive noise and the application of
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coupling theory. The second part presents a detailed
study of the structure of new energy vehicles based
on modal analysis, and the use of acoustic-solid
coupling to optimize noise control techniques for
new energy vehicles. The third part is an analysis of
the practical application of the optimization methods
proposed in the study of vehicle noise control
technology. The final section summarises and
analyses the research content.
2 Literature Review
With the increasing demand for comfort during the
driving process, many scholars have been gradually
deepening their research on automotive noise. Kim
and Cho proposed a potential semantic control
generative adversarial network method to address
the problem that it is difficult to accurately classify
noise types in automotive quality assessment. After
comparative experimental analysis, the results show
that the method has an accuracy of 96.68% in
classifying noise data, and an accuracy of 94.68%
for the interference case, [6]. Yu’s team proposes a
discrete wavelet-based noise suppression method for
automotive driveline noise to address the problem of
low efficiency of traditional methods of automotive
driveline noise suppression, [7]. [8], proposed a fan
blade with a ridged surface based on the bionic
principle to replace the conventional fan blade and
found that the sound pressure level of the fan blade
was reduced by 3.83 dB(A) compared with the
original blade after experimental analysis. [9],
proposed a method for solving the acoustic
perturbation equation based on incompressible
separated vortex simulations of convective pressure
pulsations, and after experiments, the results showed
that the method could accurately calculate the sound
pressure pulsations in the vehicle side windows, and
thus calculate the noise generated to the interior of
the vehicle. [10], proposed an improved denoise
method based on EEMD and optimal wavelet
threshold for model building of OPAX. The results
show that the noise reduction effect of this method
is better than that of the traditional method, which
can better improve the accuracy of the road
dynamics analysis model.
After more than a decade of development,
coupled solutions have now become a popular
research topic. [11], proposed a sequential iterative
fluid-solid coupling scheme for thermal analysis to
address the problem of low accuracy in loss
calculation and temperature rise prediction in high-
speed high-power switched reluctance motors, and
after empirical analysis, the results show that the
scheme is faster and more accurate than the
traditional one. Hou’s team proposed a fluid-solid
coupling-based finite element software modal
analysis method for satellite flexo. The results show
that the method can reduce the inherent frequency of
the central impression cylinder and is feasible, [12].
Li’s team proposed a method for the dynamic
tracking of multiphase vortex flow and unclear
critical vibration wave conversion using coupled.
After comparative tests, the results show that the
method can accurately calculate the nonlinear shock
components concentrated in the frequency range of
45-50 Hz, [13]. Franci addressed the problem of
incompressible tracking of fluids in fluid-structure
interactions. A fully Lagrangian finite element
method combining fluid-structure coupling with
nodal integration is proposed, and after empirical
analysis, the results show that the method has higher
accuracy and better convergence than traditional
methods, [14].
In summary, many scholars have used many high
technologies to study automotive noise in recent
years, and coupling theory has also been applied to
several fields, but the number of scholars who have
combined these two studies is still small. This study
proposes to apply the acoustic-solid coupling theory
to the noise control of new energy vehicles, hoping
to fill the gap in this research direction through this
method and provide a new idea for the subsequent
vehicle noise reduction research.
3 Research on Noise Control
Technology of New Energy Vehicles
Based on Acoustic-solid Coupling
Theory
With the continuous development of automotive
technology and the popularity of the theory of
sustainable development of resources, new energy
vehicles have gradually integrated into people’s
lives. However, when using new energy vehicles,
the noise inside the vehicle is high, which has an
impact on the comfort of the driver as well as the
occupants. Therefore, this chapter of the study
proposes to optimize the noise control technology of
new energy vehicles by using the sound-solid
coupling theory.
3.1 Research on Structural Noise of New
Energy Vehicles based on Modal Analysis
Theory
The noise generated during the driving process of
new energy vehicles will not only reduce the
comfort of the car ride and affect the
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communication of the vehicle occupants but will
also reduce the driver’s ability to cope with various
emergencies to a certain extent, thus reducing the
safety of the vehicle. There are various reasons for
the production of noise in the vehicle, and the
specific production mechanism is shown in Figure
1, [15].
As can be seen from Figure 1, the transmission
of noise within a new energy vehicle can be divided
into, according to the different transmission media,
airborne sound transmission and solid sound
transmission. Airborne sound transmission refers to
the medium to high-frequency noise transmitted by
the air cavity inside the vehicle, such as tire noise,
engine vibration noise, etc. Solid sound transmission
refers to the low-frequency noise generated by the
body wall panels under the action of external
excitation, such as road excitation, engine
excitation, etc. resulting in noise generated by body
vibration, etc., [16], [17]. Noise in the car can be
divided into three categories according to the nature
of the sound waves, namely structural noise, air
noise, and cavity resonance, as shown in Figure 2.
Interior
noise
Finite element analysis of
car interior acoustics
Vibration spectrum excited by
gear meshing at different
speeds in the transmission
system
Vibration and noise caused by
engine combustion, inertia
force, exhaust noise, etc
Suspension vibration
characteristics, shock absorber
noise characteristics, tire
noise, etc
The critical speed at which
aerodynamic excitation and
noise are generated during
high-speed driving
Exciting the
vibration and
noise of the
vehicle body
Exciting the
vibration and
noise of the
vehicle body
Excitation of low-
frequency
vibration noise
Exciting high-
frequency
vibration noise
Airborne
Airborne
Airborne
Fig. 1: Mechanism of noise generation in motor vehicles
1. Uneven road surface:
Engine transmission system,
driving system, etc
1. Vibration transmission
system:
Drive system, exhaust
system, driving system, etc
1. Vehicle body vibration and
skeleton vibration
2. Car cavity resonance
3. Cavity resonance noise
1. Inlet and exhaust
noise
Fig. 2: In-vehicle noise classification
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As can be seen from Figure 2, structural noise
is noise radiated to the vehicle occupants by forced
vibrations of the body wall panel structure. Air
noise is generated because the sound of the outside
air is transmitted through the vehicle gaps to the
interior cavity when the vehicle is traveling at high
speed. Cavity resonance noise is due to the fact that
the noise is constantly reflecting itself as it is
transmitted inside the vehicle, partially canceling
and partially superimposing. Modal analysis
techniques are a method of analysis based on
vibration theory and aimed at modal parameters,
[18]. The method can be used to understand the
inherent vibration characteristics of a structure, to
analyze complex structures, and to diagnose faults
in components. It is therefore proposed to use this
method to analyze the structure of a new energy
vehicle and to provide a basis for predicting and
controlling noise in the vehicle. Structural modal
analysis uses the idea of the finite element method
to transform the actual continuous, infinite degree
of freedom structural vibration response problem of
an object into a discrete, finite degree of freedom
vibration problem, where the differential equations
of motion for a multi-degree of freedom system are
shown in Equation (1), [19].
shown in Equation (1), [19].
F t M X C X K X
(1)
Equation (1),
M
represents the total mass
matrix of the vehicle structure;
C
represents the
total damping matrix of the vehicle structure;
K
represents the total stiffness matrix of the vehicle
structure;
Ft
represents the external load
vector of the vehicle;
X
represents the
acceleration of the vehicle;
X
represents the
velocity of the vehicle;
X
represents the
displacement vector of the vehicle. Since the free
vibration can be treated as a superposition of many
simple harmonic vibratory motions, the chi-square
equation can be obtained as shown in Equation (2).
20KM
(2)
Equation (2),
denotes the inherent frequency;
which denotes the inherent vibration pattern.
Based on the acoustic wave equation, the matrix
form of the fluctuation equation without
attenuation, ignoring viscous consumption, and the
matrix with attenuated fluctuation equation when
considering the presence of acoustic damping at the
boundary can be derived, as shown in Equations (3)
and (4).
00
T
PP
e e e e e e
M P K P R X
(3)
Equation (3),
P
e
M
represents the unit fluid
mass matrix;
e
P
represents the nodal pressure
vector;
P
e
K
represents the unit fluid stiffness
matrix;
0
T
e
R
represents the unit coupling mass
matrix; and
e
X
represents the nodal displacement
component vector.
00
PP
e e e e
T
P
e e e e
M P C P
K P R X
(4)
Equation (4),
P
e
C
represents the unit fluid
damping matrix. In the case of a car using a hard
surface with no sound-absorbing material at the
boundary, the characteristic equation is calculated
as shown in Equation (5).
20
PP
e e e
K M P
(5)
When a car is used with a hard surface bounded
by sound-absorbing material, the characteristic
equation is calculated as shown in Equation (6).
20
P P P
e e e e
K j C M P
(6)
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Equation (6),
j
represents a constant. Using
numerical analysis, the acoustic resonance
frequencies and the corresponding sound pressure
distributions within the new energy vehicle can be
obtained using Equations (5) and (6) for different
boundary conditions. To obtain the inherent
acoustic mode of the cavity itself, the cavity wall
plate must be assumed to be rigid to analyze the
acoustic modalities of the noise inside the vehicle
using the above research. To describe the acoustic
mode shape visually, the concepts of
“longitudinal”, “transverse”, “vertical” and
“toroidal” are used. The first 8 orders of acoustic
inherent frequencies and mode shapes are described
as shown in Table 1.
Table 1. Frequency and vibration patterns of the
acoustic modes
Order
Natural
frequency/Hz
Vibration mode description
1
36.04
Vertical first order
2
71.56
Vertical second order
3
75.01
Horizontal first order
4
81.96
Vertical first order and
horizontal first order
5
99.91
Vertical second-order and
horizontal first-order
6
108.05
Vertical third order
7
111.97
Vertical first order
8
124.03
Vertical first order and vertical
first order
As can be seen from Table 1, the acoustic
pressure pattern of the first-order mode varies
mainly along the longitudinal direction, with very
little variation in the other directions. The second-
order acoustic mode shape also varies along the
longitudinal direction, with two nodal surfaces of
very small acoustic pressure appearing above the
longitudinal cross-section, hence the term
longitudinal second-order acoustic mode. The
third-order acoustic mode varies mainly along the
transverse direction, with a nodal plane of near zero
sound pressure in the middle of the transverse
section, and is referred to as the transverse first-
order acoustic mode. The fourth-order acoustic
mode varies mainly along the longitudinal and
transverse directions, with an acoustic pressure
nodal plane appearing in both the transverse and
longitudinal sections, and is therefore referred to as
the longitudinal first-order and transverse first-
order modes. The fifth-order acoustic mode has two
nodal surfaces in the longitudinal section and one
nodal surface in the transverse section at the same
time and is referred to as the second and transverse
first-order acoustic mode.
The sixth-order acoustic mode appears with
three nodal planes in the longitudinal direction and
is therefore referred to as the longitudinal third-
order acoustic mode. The seventh-order acoustic
mode has one nodal plane with zero sound pressure
in the vertical section and is therefore referred to as
the vertical first-order acoustic mode. The sound
pressure vibration mode of the eighth-order
acoustic mode has two nodes in the longitudinal
section and a node with zero sound pressure in the
vertical section. In the longitudinal section, the
maximum sound pressure value appears at the front
top and back top of the cavity, and this mode has
the lowest resonance frequency in the longitudinal
direction. On the vertical section, the vertical first
mode has a nodal plane where the sound pressure is
zero, that is, the sound pressure changes the most at
this position. This mode has the lowest resonance
frequency in the vertical direction, so it is called the
longitudinal first order and the vertical first order.
The existence of vertical first-order mode and
vertical first-order mode is due to the limitation of
space structure and boundary conditions, which
makes the sound waveform a specific node and
anti-node distribution in a specific direction. The
presence of this vibration state is important for
acoustic design and control and can provide
reference and guidance when designing acoustic
systems or reducing noise. In addition, the study of
vertical and vertical modes is also helpful to
understand the propagation law of sound waves in
space, to better understand acoustic phenomena and
applications.
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3.2 Noise Control Technology for New
Energy Vehicles based on Acoustic-solid
Coupling Theory
During the driving process of a new energy vehicle,
the vibration of the body panels by external
excitation will produce a kind of pressure on the air
around the body panels inside the vehicle, thus
changing the sound pressure inside the vehicle, and
the change of the sound pressure inside the vehicle
will more or less amplify or suppress the vibration
of the body panels. This interaction between the
structure and the air forms an acoustic-solid
coupling system. The new energy vehicle is in fluid
during travel and is subjected to fluid pressure, so
the vibration equation for a vehicle structure that
considers the loading effect of fluid pressure is
shown in Equation (7).
pr
e e e e e e e e
M X C X K X F F
(7)
Equation (7),
pr
e
F
represents the fluid
pressure load vector, which is calculated as shown
in Equation (8).
'pr
e
S
F N P n dS
(8)
Equation (8),
S
denotes the interface area;
'
N
denotes the displacement unit shape function;
n
and denotes the unit normal of the interface.
Substituting the finite unit form function for the
spatial variation of pressure into Equation (8), the
structural vibration finite element equation for the
interface pressure vector can be derived as shown
in Equation (9).
e e e e e e e e e
M X C X K X R P F
(9)
Equation (9),
e
R
denotes the coupling mass
matrix. This leads to the finite element matrix
equation for the fluid-structure coupling problem,
as shown in Eq. (10).
[ ] [0] [ ] [0]
[ ] [ ] [0] [ ]
[ ] [ ]
0
[0] [ ]
ee
ee
fs P P
ee
ee
fs ee
e
Pe
e
MC
XX
M M C
PP
XF
KK
P
K
(10)
Equation (10),
[]
fs
M
denotes the cell coupling
mass matrix;
[]
fs
K
denotes the matrix with
negative signs for all elements in the coupling
matrix. The finite element analysis results of the
internal sound field of a vehicle will be influenced
by the finite element model of the vehicle body and
its internal cavity, and the accuracy of the finite
element model of the vehicle body will also be
reflected in the finite element analysis results.
Therefore, when building the finite element model
of the car body, the model needs to be simplified to
make the computer simulation model as close as
possible to the actual model of the car. Therefore,
the body model is simplified and key points are
positioned to create the body model. As each unit
node has six degrees of freedom in each direction,
i.e. three degrees of freedom for translation in the
z-axis, y-axis, and z-axis and three degrees of
freedom for rotation around the axis, the type of
unit structure designed for the study is shown in
Figure 3, [20].
~
~
z
y
x
x
y
z
xij
yij
zij
Ⅰ
Ⅱ
Ⅲ
Ⅳ
Ⅴ
Ⅵ
Fig. 3: Cell structure diagram of the degrees of
freedom
As can be seen from Figure 3, when the fluid
unit is located on the contact surface of the
acoustic-solid coupling in the interior cavity of the
vehicle, the unit node has four degrees of freedom,
including the x-axis, y-axis, and z-axis translational
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degrees of freedom, and the acoustic pressure
degrees of freedom of the sound field. When the
fluid unit is not located on the contact surface of
the air-solid in the interior cavity of the vehicle, it
has only one degree of freedom, the sound pressure
degree of freedom. The calculation of the weight
integral of the acoustic-solid coupling in the sound
field is shown in Equation (11).
22
0
( ( , , ) ( , , ) ( , , )) 0
V
p p x y z k p x y z j q x y z dV
(11)
Equation (11),
p
represents the weight
coefficient;
2
represents the Lagrangian operator;
( , , )p x y z
represents the sound pressure function
p
in the coordinate system;
k
and
q
both represent
constants; and
0
represents density. The
mathematical expression for the wavelet amplitude
in the confined compartment at this point is shown
in Equation (12).
2
022
0
1 ( , , )p x y z
Ct
(12)
Equation (12),
0
C
represents the speed of
propagation of the acoustic wave;
t
represents the
time;
represents the partial derivative. The
equation for the acoustic material state is shown in
Equation (13).
' ' ' 2
0
2
00
( 1) ( ) ...
2
p
p
(13)
Equation (13),
represents the ratio of the
constant pressure specific heat capacity to the heat
capacity of the gas;
'
represents the density of the
fluid; and
'
p
represents the acoustic pressure of the
fluid. The kinetic equation on the coordinate axis is
shown in Equation (14).
'
'0
0
'
'0
0
'
'0
0
()
( )( )
()
( )( )
()
( )( )
x
y
z
pp
vv
tx
pp
vv
ty
pp
vv
tz
(14)
Equation (14),
0
p
represents the sound pressure
at rest;
x
v
,
y
v
and
z
v
represents the velocity in the
x, y, and z directions respectively;
represents the
velocity change value. Equation (14) is combined
with the vector to obtain Equation (15).
''
00
( )( ) ( )v v p p
t
(15)
The workflow for the finite element analysis of
the acoustic-solid coupling of the noise inside the
vehicle is therefore shown in Figure 4.
Structural
model
Cavity
Model
Body cavity
model
Calculated and
measured
values of load
Noise pressure
frequency curve at
selected points and
specific operating
conditions
Contribution of
each modality
Fig. 4: Flow chart of acoustic-solid coupling
analysis
As can be seen from Figure 4, the air finite
element model of the internal cavity of the car and
the structure of the car itself, as well as the solid-
fluid coupling model of the mutual coupling
between the two, are first established. Then the
finite element simulation of the acoustic-solid
coupling is established and the response of the
sound field, sound pressure distribution, and
excitation points under a specific load condition are
calculated by the finite element simulation. The
response of the interior cavity structure of the
vehicle is obtained for each order of the modal
sound field and the contribution of the
corresponding structural body members to this
response. The finite element model of the acousto-
solid coupled system is shown in Figure 5.
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Fig. 5: Acoustic-solid coupling system model
As shown in Figure 5, since the coupled mode
is formed by the interaction between the body
structure and the air, the coupled mode vibration
distribution is partly dominated by the structural
deformation and partly by the sound pressure
variation, which corresponds to the two systematic
modes of the structure and the acoustic cavity
respectively. In summary, this paper proposes to
use the sound-structure coupling theory and the
structural modal analysis of new energy vehicles to
build a sound-structure coupling system model for
accurate monitoring of vehicle noise. Firstly, the
finite element model of the new energy body
structure is established, the modal natural
frequency and vibration mode of the body structure
are calculated, and the vibration components are
found through the structural modes. Then, the
acoustic finite element model of the interior cavity
and the acoustic-solid coupling model are
established, the acoustic natural mode and
frequency are calculated, and the sound pressure
deformation is the main mode and frequency under
the coupling action, to realize the accurate
monitoring of the interior noise of new energy
vehicles.
4 Analysis of the Application Effect of
Noise Control Technology of New Energy
Vehicles Combined with Sound-solid
Coupling
To analyze the effect of the practical application of
the research-proposed noise control technology for
new energy vehicles optimized using acoustic-solid
coupling theory, the research will use real vehicle
road tests. A new energy vehicle of a certain brand
will be used as the test vehicle, in which
acceleration sensors, microphones, and data
collectors will be installed, and a computer will be
connected to conduct the test using LMS software.
The study will take the frequency response of the
new energy vehicle body vibration and the sound
pressure frequency response inside the vehicle as
the main test objects of this test.
4.1 New Energy Vehicle Body Vibration
Frequency Response Analysis
To analyze the frequency response of the new
energy vehicle body vibration, test points were set
up on the roof and floor of the test vehicle
respectively, and the change in vibration
acceleration at the test points was recorded through
a series of external excitation signals, and the test
results are shown in Figure 6.
020 40 60 80 100 120 140 160 180 200
0.0
0.2
0.4
0.6
0.8
1.0
1.2
1.4
1.6
1.8
Frequency/Hz
Acceleration m/s2
Before improvement
Improved
(a) Roof panel test point
020 40 60 80 100 120 140 160 180 200
0.0
0.2
0.4
0.6
0.8
1.0
1.2
1.4
1.6
1.8
Frequency/Hz
Acceleration m/s2
Before improvement
Improved
(b) Underbody test point
Fig. 6: Vibration acceleration of the test point of automobile roof and bottom plate
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010 20 30 40 50
0.00
0.02
0.04
0.06
0.08
0.10
0.12
0.14
0.16
0.18
Frequency/Hz
Acceleration m/s2
(a) Before improvement
010 20 30 40 50
0.00
0.02
0.04
0.06
0.08
0.10
0.12
0.14
0.16
0.18
Frequency/Hz
Acceleration m/s2
(b) After improvement
Front axle head
On the front spring
Rear axle head
On the rear spring
Front axle head
On the front spring
Rear axle head
On the rear spring
Fig. 7: The vibration spectrum on the front and rear axle head and spring of the car
Table 2. The vibration frequency of the automobile
engine at different speeds
Driving speed (km/h)
30
60
80
Engine speed (rpm)
1000
1300
1500
Before the
improvement
Excited
frequency (Hz)
60
75
79
Vibration
frequency of
suspension rack
(Hz)
50
62
73
After the
improvement
Excited
frequency (Hz)
50
60
70
Vibration
frequency of
suspension rack
(Hz)
43
50
62
From Figure 6(a), it can be seen that at the
external excitation frequency range of 0-200 Hz,
the vibration of the test point of the car roof is
obvious at about 20 Hz, 108 Hz, and 150 Hz. The
acceleration of the test point before the improved
noise control technology is 1.61 m/s2, 0.89 m/s2 and
1.71 m/s2 respectively. The acceleration of the test
point after the improved noise control technology is
1.48 m/s2, 0.61 m/s2, and 1.43 m/s2 respectively.
From Figure 6(b), it can be seen that when the
external excitation frequency is in the range of 0-
200 Hz, the vibration of the test point of the car
floor is obvious at 5 Hz, 35 Hz, 50 Hz, 105 Hz, 153
Hz, and 195 Hz. After the improvement of noise
control technology, the acceleration of the test
points were 1.02 m/s2, 0.59 m/s2, 0.62 m/s2, 0.67
m/s2, 0.71 m/s2, 0.33 m/s2. The test points were set
up at the front axle head, the front spring, the rear
axle head, and the rear spring, and the changes in
vibration acceleration at the test points were
recorded while the car was traveling at a speed of
50 km/h. The test results are shown in Figure 7.
As can be seen from Figure 7, the external
excitation frequency range generated during the
car’s uniform speed is 0-50 Hz, with all test points
producing peak vibration at a frequency of around
15 Hz. The peak vibration accelerations at the front
axle head, on the front spring, on the rear axle
head, and the rear spring were 0.175 m/s2, 0.158
m/s2, 0.043 m/s2, and 0.059 m/s2 respectively, when
the vehicle was not using the improved noise
control technology. After improvement, the peak
vibration accelerations were 0.080 m/s2, 0.078
m/s2, 0.020 m/s2, and 0.019 m/s2 respectively. The
excitation frequencies of the car and the suspension
vibration frequencies were recorded when the car
was driven forward at different driving speeds. The
test results are shown in Table 2.
As can be seen from Table 2, when the speed of
the car is 30 km/h, the speed of the car engine is
1000 rpm, before using the improved noise control
technology, its excitation frequency is 60 Hz and
the vibration frequency of the suspension is 50 Hz;
after using the improved noise control technology,
its excitation frequency is 50 Hz and the vibration
frequency of the suspension is 43 Hz. when the
speed of the car is 60 km/h, the speed of the car
engine is 1300 rpm, before using the improved
noise control technology, its excitation frequency is
75 Hz and the vibration frequency of the
suspension is 62 Hz. At a speed of 60 km/h, the
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engine of the car rotates at 1300 rpm, before using
the improved noise control technology, the
excitation frequency is 75 Hz and the vibration
frequency of the suspension is 62 Hz; after using
the improved noise control technology, the
excitation frequency is 60 Hz and the vibration
frequency of the suspension is 50 Hz. The
excitation frequency is 79 Hz and the suspension
vibration frequency is 73 Hz; after using the
improved noise control technology, the excitation
frequency is 70 Hz and the suspension vibration
frequency is 62 Hz. The frequency response
analysis of body vibration refers to the analysis of
the vibration generated by the vehicle during
operation to determine the main frequency and
amplitude of the vibration, understand the vibration
intensity of the vehicle under different frequencies,
determine the structural strength and comfort level
of the vehicle, and detect the fault or abnormal
situation of the vehicle. In the field of automotive
engineering, the analysis of body vibration
frequency response is an important research
direction, that is of great significance in improving
the performance and safety of vehicles. Through
the comparison of the vibration acceleration
changes of the test points of the roof and side
panels of new energy vehicles, the results show that
the vibration acceleration of the test points has
decreased after improving the noise control
technology. By comparing the vibration
acceleration peaks of the test points of the front
axle head, front spring, rear axle head, and rear
spring of new energy vehicles, the results show that
the vibration acceleration peaks of the test points
after the improvement of noise control technology
are all lower than the peak before the improvement.
By recording the exciting frequency and suspension
vibration frequency of new energy vehicles at
different driving speeds, the results show that the
exciting frequency and suspension vibration
frequency of new energy vehicles are reduced after
the improvement of noise control technology. The
above results show that the technology can reduce
vibration intensity, reduce noise, and improve
comfort.
4.2 Frequency Response Analysis of
Sound Pressure in New Energy Vehicles
To analyze the sound pressure frequency response
inside the new energy vehicle, the study set up test
points at the windscreen and side window glass of
the vehicle, and recorded the noise pressure value
at the test points under the change of engine speed,
the test results are shown in Figure 8.
From Figure 8(a), it can be seen that at the test
point at the windscreen of the car, the noise
pressure value reaches the highest when the engine
speed reaches 3500 r/min, and the noise pressure
values before and after using the improved noise
control technology are 93 dB(A) and 80 dB(A)
respectively, and the noise pressure values after
using the improved noise control technology are
lower than before the improvement. As can be seen
from Figure 8(b), before the use of the improved
noise control technology, the test point at the side
window glass of the car reached a maximum noise
pressure value of 90 dB(A) when the engine speed
reached 3500 r/min. After the use of the improved
noise control technology, the test point at the side
glass of the car reached a maximum noise pressure
of 73 dB(A) at an engine speed of 2000 r/min. The
noise pressure values after the improved noise
control technology were lower than before the
improvement. The sound pressure at the ear of the
driver, co-driver, and rear passenger was measured
and recorded at different operating conditions and
the test results are shown in Table 3.
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1000 2000 3000 4000 5000
50
60
70
80
90
Noise pressure
dB (A)
Engine speed r/min
Before improvement
Improved
(a) Change in sound pressure at the
windshield
1000 2000 3000 4000 5000
50
60
70
80
90
Noise pressure
dB (A)
Engine speed r/min
Before improvement
Improved
(b) Change in sound pressure at the
side window glass
Fig. 8: Noise value at the automotive glass
Table 3. Sound pressure value in cars
Workin
g
conditi
on
(rpm)
Driver sound
pressure/dB(
A)
Co-pilot
sound
pressure/dB(
A)
Rear
personnel
sound
pressure/dB(
A)
Befor
e
Afte
r
Befor
e
Afte
r
Befor
e
Afte
r
800
78
74
80
77
76
71
1400
74
72
76
73
72
70
2500
85
81
80
77
81
79
As can be seen from Table 3, the sound pressure
at the driver’s, co-driver, and rear passenger’s ears
was 78 dB(A), 80 dB(A), and 76 dB(A)
respectively before the use of the improved noise
control technology at 800 rpm. With the improved
noise control technology, the sound pressure at the
driver’s, co-drivers, and rear passenger’s ears is 74
dB(A), 77 dB(A), and 71 dB(A) respectively. At
1400 rpm, the sound pressure at the driver’s, co-
drivers, and rear passenger’s ears was 74 dB(A), 76
dB(A), and 72 dB(A) respectively before the use of
the improved noise control technology. With the
improved noise control technology, the sound
pressure at the driver’s, co-drivers, and rear
passenger’s ears is 72 dB(A), 73 dB(A), and 70
dB(A) respectively. At 2500 rpm, the sound
pressure at the driver’s, co-driver’s and rear
passenger’s ears were 85 dB(A), 80 dB(A), and 81
dB(A) respectively before the use of the improved
noise control technology. With the improved noise
control technology, the sound pressure at the
driver’s, co-drivers, and rear passenger’s ears was
81 dB(A), 77 dB(A), and 79 dB(A) respectively.
The sound pressure at the driver’s side of the ear in
the car was recorded while the car was driven at 50
km/h on a concrete and asphalt road and the results
are shown in Figure 9.
020 40 60 80 100 120 140 160 180
10
20
30
40
50
60
70
80
90
100
Frequency/Hz
Before improvement
Improved
(a) Cement pavement
Noise pressure dB (A)
020 40 60 80 100 120 140 160 180
10
20
30
40
50
60
70
80
90
100
Frequency/Hz
Before improvement
Improved
(b) Asphalt pavement
Noise pressure dB (A)
Fig. 9: Sound pressure value curve of vehicles on different road surfaces
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Table 4. Simulation error of the finite element model
for improved noise control techniques
Serial
number
Error type
35 Hz
70 Hz
100
Hz
1
Model
assumption error
1.00%
1.04%
0.09%
2
Calculation
method error
1.01%
1.01%
1.00%
3
Calculation
accuracy error
0.09%
1.00%
1.03%
Figure 9(a) shows that when the car is driven
on concrete roads, the maximum sound pressure
next to the driver’s ear is 62 dB(A) after using the
improved noise control technology, which is lower
than 86 dB(A) before using the technology. Figure
9(b) shows that when the vehicle is driven on
asphalt, the maximum sound pressure next to the
driver’s ear is 77 dB(A) after using the improved
noise control technology, which is lower than the 92
dB(A) before using the improved noise control
technology. The frequency response analysis of the
sound pressure inside the car refers to the frequency
response analysis of the noise inside the car to
understand the amplitude and characteristics of the
noise inside the car at different frequencies. In-
vehicle sound pressure frequency response analysis
can also help detect vehicle noise problems, identify
noise sources, and find solutions. Test points were
set at the windscreen and side window of new
energy vehicles to record the noise pressure value at
the test point. The test results show that the noise
pressure value after using the improved noise
control technology is lower than before. By
detecting and recording the sound pressure near the
ear of the driver, co-driver, and rear passenger of the
new energy vehicle, we can see by comparing the
structure that under different vehicle working
conditions, the sound pressure near the ear of the
driver, co-driver and rear passenger is reduced after
using the improved noise control technology. The
results show that the maximum and minimum of the
sound pressure near the driver's ear are reduced by
using the improved noise control technology.
Combining all the above test results, the improved
noise control technique is feasible. To further
explore the performance of the improved noise
control technology, the simulation error analysis of
the finite element model is studied, and the test
results are shown in Table 4.
As can be seen from Table 4, in the simulation
error analysis of the finite element model of the
improved noise control technology, it is found that
when the internal noise frequency of new energy
vehicles is 35 Hz, 70 Hz, and 100 Hz, the model
hypothesis error, calculation method error, and
calculation accuracy error of the improved noise
control technology are all floating around 1.00%,
which is a small error. The result further verifies the
superiority of the improved noise control
technology.
At present, the commonly used internal noise
control technologies include acoustic isolation
technology, shock absorbers, active noise control
technology, engine sound insulation, aerodynamic
optimization technology, and so on. Acoustic
isolation technology can effectively isolate the
transmission and radiation of noise and provide a
good driving environment. However, this
technology increases the weight and cost of the
vehicle, which can affect the fuel economy of the
vehicle. The installation of damping devices can
effectively reduce the vibration and noise caused by
road noise and bumps, and improve driving comfort.
However, it increases the cost of the vehicle and
puts forward higher requirements for the design and
maintenance of the suspension system. Active noise
control technology uses sensors and control systems
to control reverse sound waves to counteract noise.
This technology can create offset sound waves
inside the car, reducing noise levels. The advantage
is that it can provide a better noise control effect,
which is not limited by the structure and material of
the vehicle. However, the technology requires
complex sensors and control systems, adding
complexity and cost to the vehicle. Engine
soundproofing can effectively reduce the conduction
and radiation of engine noise, providing a quieter
driving experience. However, it will have a certain
impact on the heat dissipation of the engine.
Aerodynamic optimization technology reduces noise
caused by wind noise and wind resistance by
optimizing the body and parts design but affects the
exterior design and performance of the vehicle. The
noise control technology of new energy vehicles
proposed in this study is based on the vehicle
structure of finished vehicles, and the acoustic-solid
coupling theory is proposed to carry out modal
analysis of new energy vehicle structure, optimize
the noise control technology of new energy vehicles,
and build an acoustic-solid coupling system model
to accurately monitor the internal noise of vehicles.
This method can effectively control the noise in the
driving process of new energy vehicles, reduce the
negative impact of noise, and does not increase the
cost of vehicle design and research and
development, which is better than the commonly
used internal vehicle noise control technology.
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5 Conclusion
With the continuous development of automotive
technology, the performance of cars is gradually
improved, and due to the influence of the concept of
sustainable development, people’s requirements for
cars, in addition to meeting the basic conditions of
safety performance, also require its low-carbon
environmental protection. Therefore, the emergence
of new energy vehicles has attracted much attention,
but the noise generated by these vehicles during the
driving process is large and affects the comfort of
the occupants and the concentration of the driver.
The study therefore proposes to improve the noise
control technology of new energy vehicles using
acoustic-solid coupling theory. The results show that
the maximum vibration acceleration at the top plate,
bottom plate, axle head, and spring of the vehicle
are 1.48 m/s2, 1.02 m/s2, 0.079 m/s2, and 0.020 m/s2
respectively, which are lower than the maximum
vibration acceleration before the use of the
technology. The vibration frequencies of the
suspension after the use of this technology were 43
Hz, 50 Hz, and 62 Hz respectively at different
speeds of the car, all lower than before the use. The
maximum noise pressure values detected at the
windscreen and side window glass were 80 dB(A)
and 73 dB(A) respectively after the use of this
technology. The sound pressure at the driver’s,
passenger’s, and rear passenger’s ears at 800 rpm
was 74 dB(A), 77 dB(A) and 71 dB(A) respectively,
all lower than before. The maximum sound pressure
at the driver’s ear after using the improved
technology is 62 dB(A) and 77 dB(A) respectively
when the car is driven on different roads, which are
lower than before. In summary, the study proposes
to improve the noise control technology of new
energy vehicles by combining sound-solid coupling
theory, which can effectively reduce the noise
generated by new energy vehicles and improve the
comfort of vehicle occupants. However, in the
process of research, the pores in the carriage of new
energy vehicles and the vibration of the body skin,
as well as the weight of the engine, clutch, and other
transmission systems and the air conditioning in the
car are ignored. It is hoped that we can pay more
attention to the leakage noise, strengthen the
modeling accuracy, and improve the credibility of
the simulation. At the same time, interior materials
such as seats, instrument panel assembly, and carpet
make the distribution of the interior sound field
more complicated, which has not been explored in
this study. It is hoped that the sound-absorbing
characteristics of interior materials can be corrected
by testing in the future, and the influence of interior
materials on the interior sound field can be analyzed
in detail.
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Contribution of Individual Authors to the
Creation of a Scientific Article (Ghostwriting
Policy)
The author FUJUN MAO contributed to the present
research, at all stages from the formulation of the
problem to the final findings and solution.
Sources of Funding for Research Presented in a
Scientific Article or Scientific Article Itself
No funding was received for conducting this study.
Conflict of Interest
The author has no conflicts of interest to declare that
are relevant to the content of this article.
Creative Commons Attribution License 4.0
(Attribution 4.0 International, CC BY 4.0)
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