JP3399732B2 - Method of reducing radiation noise of vehicle body panel and vehicle body panel - Google Patents

Description

Detailed Description of the Invention

[0001]

The present invention relates to relates to a vehicle body panel sound radiation reduction method and a vehicle body panel of a vehicle, in particular sound radiation effective body panels reduction method and the method to reduce the road noise during the running of the vehicle The present invention relates to a vehicle body panel with reduced radiated sound.

[0002]

2. Description of the Related Art In automobiles, in order to reduce vibrations and noises (road noises) transmitted to the vehicle interior due to resonance of suspension devices and cavity resonance of tires during traveling, various vibration control systems have been conventionally used. , Soundproofing measures are taken.

Particularly, a damping material is attached to a vehicle body panel such as a floor panel forming the floor surface of a vehicle compartment, or an actual flat panel 4-435.
As disclosed in Japanese Patent No. 78, a plurality of ribs are formed so as to intersect with each other to increase the surface rigidity of the vehicle body panel, thereby reducing vibration and noise.

[0004]

However, for example, in the case of a panel having a rib and having a high surface rigidity, the vibration level is lower than that in the case where the rib is not provided, but the sound tends to be emitted more. Therefore, even if the vibration transmitted to the vehicle interior is reduced, the noise may not be sufficiently reduced. It has also been found that if the surface rigidity is high, the damping effect of the damping material is weakened.

On the other hand, on a flat panel surface having low surface rigidity,
Many vibration modes are excited by the divided vibration,
It has also been found that there is a large difference in radiation efficiency between the vibration modes.

[0006] The present invention aims at providing a vehicle body panel sound radiation reduction method and vehicle body panels vehicles focused on the radiation efficiency of each of the vibration modes excited in the panel surface.

[0007]

SUMMARY OF THE INVENTION A method for reducing radiated sound of a vehicle body panel of a vehicle according to the present invention sets a target frequency f _S for reducing the radiated sound, and vibrates the vehicle body panel of the vehicle. The mode is an n × m mode (n = 1, 2, 3, ... Number of antinodes of vibration in a predetermined direction; m = 1, 2, 3, ... Number of antinodes of vibration in a direction perpendicular to the predetermined direction) In this case, the structural parameters of the vehicle body panel are adjusted so that the frequency of the vibration mode in which the radiation efficiency is low, n × m = even, approaches the target frequency f _S (claim 1).

When adjusting the structural parameters of the vehicle body panel, as shown in FIG. 1, (a) the target frequency f _S is set from the frequency characteristics of noise, and (b) the dimensions, materials, supporting conditions, etc. of the vehicle body panel. (C) detecting the frequency and strain energy distribution of each of a plurality of vibration modes generated in the vehicle body panel under the constraint condition, and (d) FIG. 2 (a), (b) and 3 (a), 3 (b)
To, with respect to the target frequency f _S, as shown as an example for the case you want to excite the vibration of the 2 × 2 mode low radiation efficiency, the magnitude relationship between the target frequency f _S and the frequency of each vibration mode, change 4 (a) to FIG. 4 (a), which are structural energy parameters to be determined, and (e) a strain energy distribution diagram by the finite element method in which a portion having a large strain energy is shown by a black portion.
As shown in (c), the portion where the structural parameter should be changed (the portion where the strain energy distribution is large) may be determined from the strain energy distribution of each vibration mode (claim 2).

When the vehicle body panel is not a square or a rectangle close to a square, it is preferable to form a square or a rectangle close to a square whose periphery is bound by a frame or a bead in the vehicle body panel (claim 3).

The resonance frequency of the vehicle suspension may be set as the target frequency f _S (claim 4). Specifically, when the vehicle is provided with a strut suspension, the lean resonance frequency of the strut suspension in the vehicle width direction may be set as the target frequency f _S (claim 5). Alternatively, the tire cavity resonance frequency of the vehicle may be set as the target frequency f _S (claim 6). Further, both the fall resonance frequency of the suspension in the vehicle width direction and the tire cavity resonance frequency may be set as the target frequency f _S (claim 6). The vehicle body panel is, for example, a floor panel (claim 7).

The predetermined vibration mode frequency is set to the target frequency f _S
The change of the structural parameter for approaching is performed by changing the surface rigidity distribution of the vehicle body panel (claim 8).

For example, in order to bring the frequency of the vibration mode with low radiation efficiency close to the target frequency f _S higher than that frequency, the surface rigidity is adjusted in the portion where the strain energy distribution in the vibration mode is large (claim 8). ).

Specifically, as shown in FIG. 2 (a), the surface curvature of that portion is adjusted, an out-of-plane protruding portion is formed in that portion, and the protruding amount of this protruding portion is increased. The surface rigidity is increased by performing at least one of forming the bead on the portion and forming the bead on the portion (claim 9).

Further, for example, in order to bring the frequency of the vibration mode having low radiation efficiency close to the target frequency f _S lower than that frequency, as shown in FIG. 3A, the strain energy distribution of the vibration mode is changed. The surface density of the large portion may be increased (claims 10 and 11), but it is also possible to reduce the surface rigidity of the portion by removing the beads.

Further, when the mode frequency is equal to the target frequency f _S , as shown in FIG. 3B, in order to suppress the other modes (1 × 1, 1 × 3) having high radiation efficiency, the mode is suppressed. A damping material may be attached to the portion having a large strain energy distribution (claim 11).

The vibration mode with low radiation efficiency is, for example, 2 ×
There are two modes (claim 12).

The vehicle body panel of a vehicle according to the present invention, substantially is rectangular, its vibration modes n × excited in the vehicle body panel m mode (n = 1, 2, 3, ... vibration in the predetermined direction is constrained around The number of antinodes of m; 1, 2, 3, ..., The number of antinodes of vibration in a direction perpendicular to the predetermined direction) The structural parameters are set so that the vibration modes of which the number is an even number are excited (claim 13).

[0018] The vibration frequency to be reduced over the Kiho ions, car
The resonance frequencies of both suspensions are preferably set (claim 14). Alternatively, the vibration frequency for reducing the radiated sound may be set to the tire cavity resonance frequency of the vehicle (claim 15). Furthermore, the vibration frequency that should reduce the radiated sound is the resonance frequency of the suspension,
It may be set to both the tire cavity resonance frequency (Claim 15).

The vehicle body panel is, for example, a floor panel (claim 16), and the floor panel is constrained by a vehicle body frame, side sills and a floor tunnel forming member (claim 17).

In such a vehicle body panel, as shown in FIG.
As shown in (a) and (b), the structural parameter can be set by adjusting the surface rigidity distribution of the portion of the vehicle body panel where the strain energy distribution is large (claim 18). At least one of adjusting the surface curvature of the portion, forming an out-of-plane protrusion on the portion, adjusting the protrusion amount of the protrusion, and forming a bead on the portion. By doing so, the surface rigidity is adjusted (claim 19).

Further, in this vehicle body panel, FIG.
As shown in (a), the structural parameters can be set also by adjusting the areal density distribution of the portion of the vehicle body panel where the strain energy distribution is large (claim 20). Adjustment or by attaching a damping material (claims 21 and 22).

[0022]

[Effect of the Invention] Figure 5 is a car substantially square constrained around
It is explanatory drawing which shows the relationship between the vibration mode excited in the body panel surface, and radiated sound energy.

In the 1 × 1 mode shown in FIG. 5A, the radiation efficiency is high and the radiated sound is loud, but the 1 × 2 mode shown in FIG.
In the mode, the radiated sound is canceled out, and the radiation efficiency is reduced. In addition, the 2 × 2 mode shown in FIG.
Even in the 2 × 3 mode shown in (d), the radiated sound is canceled out,
The radiation efficiency is low.

That is, in general, the vibration mode is n × m
Then, as shown in FIG. 6, n × m = even number (n = m
1, 2, 3, ... Number of antinodes of vibration in a predetermined direction; m = 1, 2,
3, ... Number of antinodes of vibration in a direction perpendicular to the predetermined direction)
If the vibration mode to be used is a panel structure in which the radiated sound is excited at a frequency at which the radiated sound should be reduced, harmful radiated sound is reduced.

FIG. 7 shows about 600 mm × 6 with the periphery constrained.
It is a graph which shows the radiation efficiency of the low-order vibration mode excited by the thin vehicle body panel of the size of 00 mm. The horizontal axis indicates the frequency ratio (kb the reference frequency), k _b / k of the vehicle body vibration is about 0.1. Also, the radiation efficiency on the vertical axis decreases toward the bottom of the figure. Referring to FIG. 7, it can be seen that when n × m = odd, the radiation efficiency is high, and when n × m = even, the radiation efficiency is low. The radiation efficiency of the 1 × 1 mode is highest even when n × m = odd, and the radiation efficiency of the 2 × 2 mode is lowest even when n × m = even.
The radiation efficiency of 2 modes is 1/1000 of the radiation efficiency of 1 × 1 mode
It can be seen that it drops to about 0.

Therefore, if the panel structure is such that the vibration mode having low radiation efficiency (1 × 2 mode or 2 × 2 mode) is excited at the frequency at which radiation noise should be reduced, harmful radiation noise is reduced. It will be.

The resonance frequency of the suspension in the vehicle both during running is transmitted to the vehicle body panel, and resulted interior noise to the vicinity of 160Hz and the peak level. Therefore, the vehicle interior noise level can be reduced by setting the resonance frequency of the suspension as the target frequency f _S.

Further, it is assumed that the tire cavity resonance caused by the standing wave generated in the tire during traveling of the traveling vehicle is transmitted to the floor panel, and the vehicle interior noise having a peak level near 250 Hz is generated. Therefore, the vehicle interior noise level can be reduced by setting the cavity resonance frequency of the tire as the target frequency f _S.

[0029]

BEST MODE FOR CARRYING OUT THE INVENTION Embodiments of the present invention will be described below.

FIG. 8 shows a square floor panel provided with a radiation noise reducing method using the method for reducing a radiation noise of a vehicle body panel according to the present invention. (Mode control panel), and the panel before countermeasures are simply called P) is a perspective view showing the front floor of a vehicle equipped with the left and right sides of the floor tunnel. Left and right edges are restrained by 10 and side sill 11, and No. 2 cross member 12 and No.
Since the front and rear edges are constrained by the three cross members 13, the entire circumference is constrained.

FIGS. 9 and 10 show No. 2 of the four floor panels MCP provided with the radiation noise reduction measures according to the present invention.
FIG. 3 is a perspective view showing a floor of a vehicle when it is provided in front of and behind a cross member 12, and the floor tunnel MCP side of a floor panel MCP is constrained by a tunnel side drain force and is constrained around the entire circumference thereof. . The four circles drawn on each floor panel MCP indicate a convex portion or a concave portion for a radiation noise reduction measure as described later, and in FIGS. 9 and 10, the convex portion on the front floor panel MCP is shown. Or the arrangement of the recesses is different.

Generally, the floor panel is provided with a hole for draining water, a hole for positioning, or a projection, but these positions adversely affect the operation of the MCP (for example, 2
The arrangement of the convex portions or concave portions is determined so that the strain energy of the × 2 mode is not small.

11 and 12 show a vehicle according to the present invention.
6 is a flowchart for explaining an embodiment in the case where the mode frequency of the vibration mode (1 × 2 mode or 2 × 2 mode) having low radiation efficiency is brought close to the target frequency f _S in the method for reducing the radiation sound of the vehicle body panel. Hereinafter, description will be given along this flowchart. In addition, S represents each step.

<Step 1> Setting of Target Frequency f _S for Reducing Radiated Sound: In the method of the present invention, first, the frequency for reducing radiated sound is set as the target frequency f _S.

As a result of frequency analysis of general vehicle interior noise,
When the vehicle has a strut suspension, the tilt resonance in the vehicle width direction of the strut suspension during traveling of the vehicle is transmitted to the floor panel P, and as shown in FIG. 13, a 1/3 octave band analysis is performed. Is performed, a noise level peak is generated in the 160 Hz band. Therefore, 160 Hz is set as the target frequency f _S for reducing the radiated sound. Since this octave band analysis is generally known, a description thereof will be omitted here (for example, Koji Kitamura, “New Edition: System Measurement of Noise and Vibration”, published by Corona Publishing Co., Ltd., February 10, 1975, first edition. See Issue). The unit "band" is generally used for sound measurement and analysis, and the word "band" is also described as "band". The “vibration frequency” in claims 16 to 18 of the present specification includes a value having a width of this “band”.

In FIG. 13, the peak of the noise level is recognized even in the 250 Hz band, which is the frequency of the cavity resonance of the tire due to the standing wave generated in the tire during traveling of the traveling vehicle. When the level is also high, it will be 250 in addition to 160 Hz, as will be described later.
Hz may also be set as the target frequency f _S.

The resonance frequency of the suspension system can be measured by using a measuring device schematically shown in FIG.
That is, vibration in the vehicle width direction is applied from the vibration exciter 2 to the wheel of the vehicle body 1 shown in the plan view of FIG.
And the output of the acceleration sensor 3 attached to the upper part of the suspension can be input to the frequency analyzer 4 to measure the resonance frequency.

The cavity resonance frequency of the tire is shown in FIG.
It can be measured using a measuring device schematically shown in. That is, vertical vibration is applied to the tire 5 from the exciter 2, and the output of the exciter 2 and the output of the acceleration sensor 3 mounted on the upper portion of the suspension are input to the frequency analyzer 4 to cause resonance. The frequency can be measured.

<Step 2> Extraction of Constraint Conditions: Next, the constraint conditions such as the size, material and support condition of the panel P are extracted. The initial conditions of panel P are as shown in FIG.
The flat panel made of SPC has dimensions of 450 × 450 × 0.65 mm (length × width × thickness) and is constrained by the frame F on the entire circumference.

<Steps 3 and 4> Check of panel shape: It is checked whether the shape of the panel P is substantially square (or rectangular) (S3), and as described above, the shape of the square (or a rectangle close to the square) is checked. In that case, it can be left as is (S
3: YES), square (or rectangle close to square)
If not (S3: NO), a square panel (or a rectangle close to a square) whose periphery is constrained by a frame or a bead is formed in the panel P, and a radiation noise reduction measure is applied to this panel (S4). The reason for this is that the 2x2 mode obtained with a square panel has a lower radiation efficiency.

<Step 5> Frequency survey and strain energy distribution survey of each vibration mode of the panel P having the initial shape: The frequency and strain energy distribution of each vibration mode of the panel P can be obtained by calculation analysis. And
2 (a) and 2 (b), which are countermeasures when it is desired to excite a 2 × 2 mode having a relatively low radiation efficiency with respect to a target frequency in a later step from the result of investigation of the frequency of each vibration mode.
As shown in FIGS. 3A and 3B, the change parameter is determined, and the strain energy distribution (the strain energy in the black portion is large) by the finite element method shown in FIGS. 4A to 4C is determined. The change position is determined from the survey results.

FIG. 17 is a schematic diagram when the vibration mode frequency of the panel P (single unit) is actually measured by using a measuring device. Floor panel P cut out from the car body 1 in the soundproof box 6
Is attached with the surroundings constrained, and a vertical vibration is applied to the panel P from the vibrator 2, and the output of the vibrator 2 and the output of the acceleration sensor 3 mounted on the floor panel P are frequency-analyzed. The sound intensity probe 7 is provided in the vicinity of the upper surface of the panel P, and the output of this probe 7 is input to the sound intensity analyzer (A.I.I.
In addition to the analyzer 8), the outputs of both analyzers 4 and 8 can be input to the computer 9 to determine the frequency, radiated acoustic power and radiation efficiency of each vibration mode.

FIG. 18 is a schematic view of a vibration mode frequency measuring device for a floor panel of an actual vehicle. In this case, for example, vertical vibration is applied from the vibrator 2 to the No. 2 cross member 12 (see FIGS. 9 and 10) of the vehicle body 1 and the output of the vibrator 2 and the floor panel P are applied. The output of the attached acceleration sensor 3 and the frequency analyzer 4 are input, and an acoustic intensity probe 7 is provided near the upper surface of the floor panel P. The output of the probe 7 is added to the acoustic intensity analyzer 8 to perform both analyzes. The outputs of the instruments 4 and 8 can be input to the computer 9 to determine the frequency, radiated acoustic power and radiation efficiency of each vibration mode.

<Steps 6 and 7> Comparison between the frequency f _{n of the} vibration mode with low radiation efficiency and the target frequency f _S : of the vibration mode with low radiation efficiency (1 × 2 mode or 2 × 2 mode) in the flat panel P The frequency f _n is the target frequency f _S
(S6), if f _n > f _S (S6: YES), the process proceeds to S8, and if f _n ≦ f _S (S6).
6: NO), it is determined whether or not f _n is lower than the target frequency f _S (S7), and if f _n <f _S (S7: YES), it is shown in FIGS. 2 (a) and 2 (b). As described above, in order to bring f _n closer to f _S by changing the surface rigidity distribution, the process proceeds to S9 in FIG. 12, and f _n =
If f _S (S7: NO), the process proceeds to S13 of FIG.

<Step 8> Selection of structural parameter to be changed: When f _n > f _S (S6: YES), the structural parameter is changed from the constraint condition of the panel P, that is, the dimension, the material, the supporting condition, and the like. It is selected whether to change the surface rigidity distribution (S9) or change the surface density distribution (S11).

More specifically, when f _n > f _S , when f _n is lowered to match f _S , (1) the surface rigidity distribution is changed, that is, the strain energy in FIG. 4A is large. If there are beads or unevenness in the black area, the black area is removed to flatten the fn.

(2) Modification of the areal density distribution, that is, even if there are beads or uneven shapes in the black area where the strain energy is large in FIG. Does not affect rigidity,
In addition, attach materials that have a large effect of adding weight. In order to reduce the rigidity of the flat panel, it is technically difficult because a part of the flat panel has to be thinned, and there is no choice but to change the areal density distribution.

<Steps 9 and 10> Adjustment of f _n by Changing Surface Stiffness Distribution: (Example 1) Adjustment of f _n by Formation of Curved Surface As a result of the investigation in step 5, excitation in a panel (flat panel) in the initial state If a large portion of the distortion energy of the 1 × 2 mode exists in the region shown by the diagonal lines in FIG. 19 and the vibration frequency thereof is, for example, 56.4 Hz, as shown in FIG. By forming a curved surface in the portion to increase the surface rigidity, the vibration frequency f _n of the 1 × 2 mode can be brought close to the target frequency f _S (= 160 Hz). In addition, in FIG. 20, FIG.
P is a perspective view, FIG. (B) is a plan view, FIG. (C) is a sectional view taken along the line cc of FIG. (B), and FIG. (D) is a view of FIG. The figure and figure (e) are figures which looked at figure (b) from the direction of arrow e.

In this case, as shown in FIG. 21, the depth of the recess of the curved surface is changed to change the local rigidity, so that FIG.
As shown in 2, the vibration frequency f _n changes. In this example, the depth of the recess of the curved surface was about 4.5 mm, and the vibration frequency of the 1 × 2 mode could be 160 Hz.

(Example 2) Adjustment of f _n by curved surface and bead As a result of the investigation in step 5, the flat panel P in the initial state is
The strain energy distributions of the 2 × 2 mode (solid line) and the diagonal 2 × 2 mode (dashed line) excited in the region are as shown in FIG. 23 (diagonal line), and the vibration frequency of the 2 × 2 mode is 77 Hz, The vibration frequency of the 2 × 2 mode was 53 Hz. The vehicle interior noise level is 160 as shown in FIG.
Peaks are shown in both Hz and 250 Hz bands.

In this case, as shown in FIGS. 24 and 25, for the 2 × 2 mode, the surface rigidity is increased by providing a recess with a depth of 5 mm in the portion where the strain energy is large,
For the diagonal 2 × 2 mode, the surface rigidity is increased by providing the bead 21 extending in the center direction in the portion where the strain energy is large. Note that, in FIG. 24, FIG. 24A is a perspective view of the panel MCP, FIG. 24B is a plan view, FIG. 24C is a sectional view taken along line cc in FIG. Figure (b)
Is a view as seen from the direction of arrow d, and FIG. 6E is a view as seen from the direction of arrow e of FIG.

As a result of the calculation and analysis, as shown in FIG. 26, the vibration frequency of the oblique 2 × 2 mode which produces an antinode on the center line is 18
The vibration frequency increased to 4 Hz, and as shown in FIG. 27, the vibration frequency of the 2 × 2 mode in which an antinode was formed on the diagonal line increased to 274 Hz. As shown in FIG. 28, the noise level in the passenger compartment during steady running at a speed of 60 km on a coarse-grained road could be reduced by about 5 dB.

(Example 3) Calculation analysis when four convex surfaces with a height of 2 mm are arranged on the surface P of the adjustment panel P of f _n by the arrangement of four convex portions (or concave portions) as shown in FIG. As a result, as shown in FIG. 30, the vibration frequency of the 2 × 2 mode that causes an antinode on the diagonal line was 172 Hz.

(Example 4) Adjustment of f _n by two convex portions and two concave portions arranged on a diagonal line As shown in FIG. 31, two diagonal lines having a height of 3 mm are arranged on one diagonal line of the panel P surface. Place the convex part on the other diagonal with a depth of 3 mm.
According to the calculation and analysis result in the case of arranging the individual concave portions, as shown in FIG. 32, the vibration frequency of the 2 × 2 mode that causes an antinode on the center line was 164 Hz.

(Example 5) Adjusting f _n by Beads As shown in FIG. 33, diagonally diagonal beads 22 are formed on the surface P of the adjustment panel.
As shown in, the vibration frequency of the 2 × 2 mode in which a diagonal antinode is 159 Hz.

<Steps 11 and 12> Adjustment of f _n by changing the areal density distribution: (Example) When a frequency lower than the 2 × 2 mode frequency of the panel P is set as the target frequency fS As shown in FIG.
When the 50 Hz band peak occurs in the vehicle interior noise level, as shown in FIG. 36, the mass 23 is added along the diagonal line to increase the surface density on the diagonal line. FIG. 37 is a diagram showing the relationship of the frequency change in the 2 × 2 mode with respect to the added weight when a heavy material is added to the panel surface in order to increase the areal density of the panel. In this case, by adding 480 g of mass, the 2 × 2 mode frequency that causes an antinode on the diagonal line as shown in FIG. 36 is reduced to 50.0 Hz, and as a result, as shown in FIG. Level is about 7 dB
Fell.

<Steps 13 and 14> Suppression of vibration modes other than the 2 × 2 mode by changing the damping distribution: When a frequency close to the 2 × 2 mode frequency of the flat panel P is set as the target frequency f _S , the vibration suppression is performed. The damping distribution is changed by attaching the material to suppress vibration modes other than the 2 × 2 mode (S13).

For example, as shown in FIG. 39, the noise level in the passenger compartment is close to the 2 × 2 mode frequency of the flat panel 80.
When peaks are shown in Hz, by attaching a damping material to the panel as shown in FIG. 3 (b), the acoustic radiation power of the panel as a whole is reduced as shown in FIG. The noise level also decreased as shown in FIG.

Further, S9, S10 and S11, S12
Even after the process of (1), it is preferable to suppress the vibration mode other than the 2 × 2 mode by changing the damping distribution by attaching the damping material (S14).

<Combination of Surface Rigidity Distribution Change and Area Density Distribution Change (for 2 × 2 mode)> (1) Surface Rigidity Distribution Change: The noise level in the passenger compartment peaks in the 160 Hz band and the 250 Hz band as shown in FIG. 24 mm, the depth of the depression in FIG. 24 described above is 5 mm.
By adjusting the curved surface, the f _n is adjusted for 160 Hz and 250 Hz.

By changing the surface rigidity distribution, the frequencies of the main modes (center frequency of the 1/3 octave band) are as shown in Table 1 below, and the 1 × 1 mode with high radiation efficiency is 10
It occurs in the 0 Hz band and has an adverse effect.

[0062]

[Table 1]

(2) Change of areal density distribution: As shown in FIG. 42, the areal density of the central portion is increased by attaching the damping material 24 to shift the frequency of the 1 × 1 mode to the low frequency range. As a result, the frequencies of the main modes (center frequency of the 1/3 octave band) are as shown in Table 2 below, and the 1 × 1 mode frequency was shifted to the 50 Hz band, and adverse effects were reduced. 42, (a) is a perspective view of the panel MCP, (b) is a plan view, and (c) is c- in FIG.
Sectional drawing which followed the c line, FIG. (d) is the figure seen from the direction of arrow d of FIG. (b), FIG. (e) is the figure seen from the direction of arrow e of FIG. (b).

[0064]

[Table 2]

<Example of design change from actual vehicle panel> (1) Change in surface rigidity: As shown in FIG. 43, the dimensions are 450 × 450 × 0.65 mm (length × width × thickness) similar to the above. A floor panel P made of SPC, in which three beads 25 in the vertical direction are provided in parallel with each other and has a high surface rigidity and whose entire circumference is restricted, is set to an initial state. Note that, in FIG. 43, FIG. 43A is a plan view of the panel P, and FIG.
A view seen from the direction of arrow b in FIG.
A cross-sectional view taken along the line, and FIG. 7D is a view seen from the direction of arrow d in FIG. In this case, as shown in Table 3 below,
Neither the 2 × 2 mode nor the diagonal 2 × 2 mode is generated, and the 1 × 1 mode exists in the 160 Hz band and the 3 × 1 mode exists in the 250 Hz band, which has an adverse effect.

[0066]

[Table 3]

Therefore, as shown in FIG. 44, the left and right beads 25 are made shallow or the inclination is made gentle to reduce the surface rigidity, and a bead 26 in the lateral direction is added along the center line of the panel P. By doing so, as shown in Table 4 below, the 1 × 1 mode is shifted to the 50 Hz band,
The 3 × 1 mode disappeared and a 2 × 2 mode having a noise reduction effect appeared in the 160 Hz band. Note that, in FIG. 44, FIG. 44A is a plan view of the panel MCP, FIG. 44B is a view seen from the direction of arrow b in FIG. 40A, and FIG. Is a cross-sectional view taken along the line, and FIG. 3D is a view seen from the direction of arrow d in FIG.

[0068]

[Table 4]

(2) Change of areal density: As shown in FIG.
It has the same dimensions of 450 × 450 × 0.65 mm (vertical × horizontal × thickness) as described above, but three vertical ribs 27 are provided in parallel, and the entire circumference having a high surface density is constrained. The SPC floor panel P is set to the initial state. Note that, in FIG. 45, FIG. 45A is a plan view of the panel P, and FIG.
Is a view as seen from the direction of arrow b in FIG. 7A, FIG. 7C is a cross-sectional view taken along line cc in FIG. 7A, and FIG.
It is the figure seen from the direction of arrow d. In this case, Table 5 below
As shown in Fig. 2, neither the 2x2 mode nor the diagonal 2x2 mode is generated, and there is a high level 1x1 mode in the 50Hz band, and the 3x1 mode is present in the 100Hz band, which causes adverse effects. ing.

[0070]

[Table 5]

Therefore, as shown in FIG. 46, the height of the left and right ribs 27 is reduced or the width thereof is reduced to reduce the surface density, and a rib 28 in the lateral direction is added along the center line of the panel P. By doing so, as shown in Table 6 below,
The level of the 1x1 mode decreases, the 3x1 mode disappears, and the 2x2 mode having a noise reduction effect has 16 levels.
Appeared in the 0 Hz band. 46, (a) is a plan view of the panel MCP, (b) is a view seen from the direction of the arrow (b) in (a), and (c) is a line c-c in (a). Is a cross-sectional view taken along the line, and FIG. 3D is a view seen from the direction of arrow d in FIG.

[0072]

[Table 6]

FIG. 47 shows a floor panel MC according to the present invention.
An enlarged perspective view showing the front floor of a vehicle equipped with P on the left and right of the floor tunnel 10 in more detail than in FIG.
48 is a sectional view taken along the line AA of FIG. The left and right edges of the two floor panels MCP are constrained by the floor tunnel 10 and the side sills 11, and the front and rear edges are constrained by No. 2 and No. 3 cross members (not shown). Then, as in FIG. 31, two convex portions 14 and 1 having a height of about 3 mm are provided on one diagonal line of the panel P surface.
4 on the other diagonal line with two recesses 1 with a depth of about 3 mm
5 and 15 are arranged, and the 2 × 2 mode that causes an antinode on the center line is excited at about 160 Hz to reduce the vehicle interior noise.

In FIG. 47, a drainage hole 16 is formed in the center of one recess 15 of both floor panels MCP, and a positioning hole 17 is formed in the other recess 15. Due to the presence of 16 and 17, the noise reduction effect of the floor panel MCP does not deteriorate.

FIG. 49 is a plan view showing an example of a floor panel MCP provided with four concave portions 15 on a diagonal line, and FIGS. 50 and 51 are lines AA and B- of FIG. 49, respectively.
It is sectional drawing along B. Further, FIG. 52 is a partial plan view of a front floor of a vehicle including a floor panel MCP provided with four recesses 15 on a diagonal line, and FIG. 53 is a sectional view taken along line AA of FIG. Is. Reference numeral 2 in FIG.
Lines indicated by 9 and 30 are bent portions provided to improve rigidity.

Further, FIG. 54 shows the front floor of a vehicle in which the floor panel MCP according to the present invention having four recesses 15 on the diagonal is provided on the left and right sides of the floor tunnel 10 in more detail than in FIG. 10 described above. It is an expansion perspective view. The line indicated by reference numeral 31 in FIG. 54 is a bent portion provided to improve rigidity.

[0077] In the above description, configuration of the vehicle body panel of a vehicle with reduced sound radiation by the radiation noise reduction method and the method of the vehicle path <br/> panel for a vehicle according to the present invention revealed,
The present invention is not only Furoapane le is adaptable to a variety of vehicle body panels that can cause vibration and noise radiation is transmitted.

[Brief description of drawings]

FIG. 1 is an explanatory diagram of processing performed on a flat panel in an initial state prior to radiation noise reduction measures in an embodiment of a method for reducing radiation noise of a vehicle body panel according to the present invention.

FIG. 2 is an explanatory diagram showing a structural parameter adjustment mode selected according to a magnitude relationship between a target frequency and a vibration mode frequency.

FIG. 3 is an explanatory diagram showing a structural parameter adjustment mode selected according to a magnitude relationship between a target frequency and a vibration mode frequency.

FIG. 4 is a strain energy distribution diagram of a vibration mode excited in a flat panel by a finite element method.

FIG. 5 is an explanatory diagram showing a relationship between a vibration mode excited in a panel surface and radiated sound energy.

FIG. 6 is a diagram showing an n × m mode excited in a panel plane.

FIG. 7 is a graph showing the relationship between vibration mode and radiated sound energy.

FIG. 8 is a perspective view showing an example of a vehicle body floor including a floor panel according to the present invention.

FIG. 9 is a perspective view showing another example of a vehicle body floor including a floor panel according to the present invention.

FIG. 10 is a perspective view showing still another example of a vehicle body floor including a floor panel according to the present invention.

FIG. 11 is a first half portion of a flowchart for explaining an embodiment of a method for reducing radiated sound of a vehicle body panel of a vehicle according to the present invention.

FIG. 12 is a latter half of a flow chart used for explaining an embodiment of a method for reducing a radiation sound of a vehicle body panel of a vehicle according to the present invention.

FIG. 13 is a frequency characteristic diagram of vehicle interior noise generated in a general vehicle.

FIG. 14 is a schematic diagram of a suspension system resonance frequency measuring device.

FIG. 15 is a schematic view of a tire cavity resonance vibration frequency measuring device.

FIG. 16 is a plan view of a flat panel to which the method for reducing the radiation sound of the vehicle body panel according to the present invention is applied.

FIG. 17 is a schematic view of a vibration mode frequency measuring device for a panel (single unit).

FIG. 18 is a schematic diagram of a vibration mode frequency measuring device for a floor panel of an actual vehicle.

FIG. 19 is a strain energy distribution map of the 1 × 2 mode excited in the panel in the initial state.

FIG. 20 is a perspective view, a trihedral view, and a cross-sectional view of a panel in which a mode frequency is adjusted by forming a curved surface on a portion of the panel surface where a large amount of strain energy is formed to enhance surface rigidity.

FIG. 21 is an explanatory diagram of a method of changing local depth by changing the depth of the depression.

FIG. 22 is a graph showing the relationship between the depth of the depression and the mode frequency.

FIG. 23 is a strain energy distribution diagram of 2 × 2 mode and oblique 2 × 2 mode excited in the panel in the initial state.

FIG. 24 is a perspective view, a three-view drawing, and a cross-sectional view of a panel in which a mode frequency is adjusted by forming a bead on a portion of the panel surface where the strain energy is large to enhance surface rigidity.

FIG. 25 is a view showing a curved surface state of the panel shown in FIG. 24.

FIG. 26 is a distortion energy distribution diagram of oblique 2 × 2 mode generated in the 160 Hz band after the radiation noise reduction measures.

FIG. 27 is a distortion energy distribution map of 2 × 2 modes generated in the 250 Hz band after the radiation noise reduction measures.

FIG. 28 is a frequency characteristic diagram of vehicle interior noise showing the noise reduction effect by changing the surface rigidity distribution.

FIG. 29 is a perspective view of a panel in which a convex surface is formed on the panel surface to change the surface rigidity distribution.

FIG. 30 is a strain energy distribution diagram of 2 × 2 mode generated in the 160 Hz band in the panel shown in FIG. 29.

FIG. 31 is a perspective view of a panel in which a convex portion is formed on the panel surface to change the surface rigidity distribution.

32 is a distortion energy distribution diagram of oblique 2 × 2 modes generated in the 160 Hz band in the panel shown in FIG. 31.

FIG. 33 is a perspective view of a panel in which beads are formed on the panel surface to change the surface rigidity distribution.

34 is a strain energy distribution diagram of 2 × 2 mode generated in the 160 Hz band in the panel shown in FIG. 33.

FIG. 35 is a frequency characteristic diagram of vehicle interior noise having a peak level at 50 Hz.

[Fig. 36] 2 × generated in the 50 Hz band after radiation noise reduction measures
Two-mode strain energy distribution map

FIG. 37 is a graph showing the relationship between load weight and mode frequency.

FIG. 38 is a frequency characteristic of vehicle interior noise showing an effect of reducing vehicle interior noise by changing the areal density distribution.

FIG. 39 is a frequency characteristic diagram of vehicle interior noise caused by panel vibration in an initial state.

FIG. 40 is a frequency characteristic diagram showing an effect of reducing acoustic radiation energy of a panel by attaching a damping material.

FIG. 41 is a frequency characteristic diagram of vehicle interior noise showing a noise reduction effect by attaching a damping material.

FIG. 42 is a perspective view, a trihedral view, and a cross-sectional view of a panel in which the mode frequency is adjusted by changing both the surface rigidity distribution and the surface density distribution of the panel.

FIG. 43 is a three-view and cross-section view of an initial panel with beads.

FIG. 44 is a three-view drawing and a cross-sectional view of a panel in which a radiation noise reduction measure according to the present invention is applied to the panel of FIG. 43.

FIG. 45 is a trihedral and cross-sectional view of an initial panel with ribs.

46 is a trihedral view and a cross-sectional view of a panel obtained by applying a radiation noise reducing measure according to the present invention to the panel of FIG. 45.

FIG. 47 is a perspective view showing an example of a front floor of a vehicle body equipped with a vehicle body panel according to the present invention.

48 is a sectional view taken along the line AA of FIG. 47.

FIG. 49 is a plan view showing an embodiment of a vehicle body panel of a vehicle according to the present invention.

50 is a sectional view taken along the line AA of FIG. 49.

51 is a cross-sectional view taken along the line BB of FIG. 49.

FIG. 52 is a plan view showing another configuration of the front floor of the vehicle body including the vehicle body panel of the present invention.

53 is a sectional view taken along the line AA of FIG. 53.

FIG. 54 is a perspective view showing still another configuration of the front floor of the vehicle body including the vehicle body panel according to the present invention.

[Explanation of symbols]

1 car body 2 shakers 3 Acceleration sensor 4 frequency analyzer 5 tires 7 Sound intensity probe 8 Sound intensity analyzer 9 personal computer 10 floor tunnel 11 side sills 12 No.2 Cross member 13 No.3 Cross member 14 convex 15 recess

─────────────────────────────────────────────────── ─── Continued Front Page (72) Inventor Norihiko Nakao 3-1, Shinchi, Fuchu-cho, Aki-gun, Hiroshima Prefecture Mazda Co., Ltd. (72) In-house Takanobu Kamura 3-1, Shinchi, Fuchu-cho, Aki-gun, Hiroshima Prefecture Da Co., Ltd. (56) Reference JP-A-6-107234 (JP, A) Actual development 4-43578 (JP, U) (58) Fields investigated (Int.Cl. ⁷ , DB name) B62D 25 / 20

Claims (22)

(57) [Claims] [Claim 1] sets a target frequency f _S should reduce radiated noise, vibration modes n × excited in the body panel of the vehicle (P) m mode (n = 1,2,3, ... in a predetermined direction The number of antinodes of vibration; m = 1, 2, 3, ... The number of antinodes of vibration in a direction perpendicular to the predetermined direction)
The frequency of the vibration mode where m = even is the target frequency f
_A method for reducing radiated sound of a vehicle body panel of a vehicle , comprising adjusting structural parameters of the panel (P) so as to approach _S. 2. (a) Extracting constraint conditions such as dimensions, materials and supporting conditions of the vehicle body panel (P), and (b) a plurality of vibration modes generated in the vehicle body panel (P) under the constraint conditions. And (c) the structural parameter to be changed is determined from the magnitude relationship between the target frequency f _S and the frequency of each vibration mode, and (d) the strain energy of each vibration mode. The method for reducing radiated sound of a vehicle body panel of a vehicle according to claim 1, characterized in that a portion where the structural parameter is to be changed is determined from a distribution. 3. If the body panel (P) is not a square or a rectangle close to a square, a square or a rectangle close to a square bounded by a frame or a bead is formed in the body panel (P). The method for reducing radiated sound of a vehicle body panel of a vehicle according to claim 2. 4. Car Vehicle according to any one of claims 1 to 3, characterized in that to set the resonance frequency of the suspension of the vehicle as the target frequency f _S
Method of reducing radiated sound from body panel. 5. A includes the vehicle strut type suspension, the vehicle according to claim 4, wherein setting the fall resonance frequency of the vehicle width direction of the strut suspension as the target frequency f _S Method of reducing radiated sound from body panels. 6. Before SL vehicle body panel <br/> Le vehicle according to any one of claims 1, characterized in that the tire cavity resonance frequency of the vehicle is set as the target frequency f _S 5 Method of reducing radiated sound. 7. The method for reducing radiated sound from a vehicle body panel of a vehicle according to claim 4, wherein the vehicle body panel (P) is a floor panel. 8. The changes of the structure parameters, the vehicle according to claim 2, wherein a carried out by changing the surface rigidity distribution of a large portion of the strain energy distribution of the vehicle body <br/> panel (P) Method for reducing radiated sound from car body panels. 9. The surface rigidity distribution of a portion of the vehicle body panel (P) having a large strain energy distribution is adjusted by adjusting surface curvature, forming protrusions (14, 15), and projecting amounts of the protrusions (14, 15). 9. The method for reducing radiated noise of a vehicle body panel according to claim 8, wherein the method is changed by at least one of adjustment and formation of a bead (22). 10. The change of the structural parameter is performed by the vehicle.
The method for reducing radiated sound of a vehicle body panel of a vehicle according to claim 2, wherein the method is performed by changing an areal density distribution of a portion of the body panel (P) having a large strain energy distribution. 11. The area density distribution of a portion of the vehicle body panel (P) having a large strain energy distribution is changed by adjusting the area density and / or attaching a damping material (24). Method of reducing radiated sound from vehicle body panel of vehicle . 12. The method for reducing radiated sound of a vehicle body panel of a vehicle according to claim 1, wherein the vibration mode is a 2 × 2 mode. 13. A vehicle body panel (P) for a vehicle, comprising :
The body panel (P) has a substantially rectangular shape with its periphery constrained.
The vibration mode excited in the vehicle body panel (P).
Xm mode (n = 1, 2, 3, ... Number of antinodes of vibration in a predetermined direction; m = 1, 2, 3, ... Number of antinodes of vibration in a direction perpendicular to the predetermined direction) A vehicle body panel, wherein structural parameters are set such that a vibration mode having a low radiation efficiency of n × m = even is excited at a vibration frequency at which radiation noise should be reduced. 14. previous SL oscillation frequency should reduce radiated noise, vehicle path <br/> panel of a vehicle according to claim 13, characterized by being set to the resonance frequency of the suspension of the vehicle. 15. Before SL oscillation frequency should reduce radiated sound, car vehicle according to claim 13 or 14, wherein the composed is set to the tire cavity resonance frequency of the vehicle
Body panel. 16. The vehicle body panel (P) according to claim 14 or 15, characterized in that it comprises a floor panel.
Vehicle body panel. 17. The vehicle body panel according to claim 16, wherein the floor panel (P) is constrained by a vehicle body frame (12, 13), a side sill 11 and a floor tunnel 10 forming member. . 18. The structural parameter is set by adjusting a surface rigidity distribution of a portion of the vehicle body panel (P) having a large strain energy distribution. Of the vehicle described in paragraph 1
Body panel. 19. Adjustment of the surface curvature of a portion of the vehicle body panel (P) where the strain energy distribution is large, and the protrusions (14, 1)
5) formation, adjustment of the amount of protrusion of the protrusions (14, 15),
19. The vehicle body panel according to claim 18, wherein the surface rigidity distribution is adjusted by at least one of the formation of the bead (22) and the formation of the bead (22). 20. The structural parameter is set by adjusting an areal density distribution of a portion of the vehicle body panel (P) having a large strain energy distribution. Of the vehicle described in paragraph 1
Body panel. By adjusting the surface density of the large portion of the strain energy distribution of claim 19, wherein the vehicle body panel (P), the vehicle body panel of a vehicle according to claim 20, wherein the comprising performed to adjust the surface density distribution. 22. The areal density distribution is adjusted by attaching a damping material (24) to a portion of the vehicle body panel (P) having a large strain energy distribution. Vehicle body panel.