Vibration-damping structure and automobile part equipped with the vibration-damping structure

Abstract

To provide a vibration control structure which can easily suppress lower-frequency vibration without increasing the mass of the vibration control structure.SOLUTION: A vibration control structure of the present invention comprises a vibration control member having a wedge-shaped part, and an elastic member. A thickness of the wedge-shaped part changes from a side on which vibration enters to a free end to which the vibration propagates satisfies the following Formula (1). The elastic member is a plate-shaped member with a Young's modulus lower than that of the vibration control member, and a part of the plate-shaped member is fixed while protruding from the free end of the wedge-shaped part. Thus, a spring constant of the elastic member can be changed via a protrusion length of the elastic member and low-frequency vibration can be easily suppressed. h(x)=ε.xn+h0...formula (1). Here, in Formula (1), x: a distance (mm) from the free end, h(x): a thickness (mm) at a distance x from the free end, h0: a thickness (mm) of the free end, ε: a positive constant, and n: a real number equal to or larger than 1.SELECTED DRAWING: Figure 2

Description

[Technical Field]

[0001] The present invention relates to a vibration-damping structure and an automobile part equipped with the vibration-damping structure, and more particularly to a vibration-damping structure that utilizes the acoustic black hole effect and an automobile part equipped with the vibration-damping structure. [Background technology]

[0002] As a countermeasure against vibration and noise, it is known to use a vibration-damping structure that utilizes the acoustic black hole effect.

[0003] This vibration-damping structure has a wedge-shaped structure in which the plate thickness decreases toward the tip.The vibrations (waves) propagating through this vibration-damping structure increase in amplitude and their propagation speed slows as they approach the tip of the wedge.At the tip of the wedge, where the plate thickness becomes 0 (zero), the propagation speed becomes 0 and the vibrations are not reflected.This is what is used.

[0004] However, it is not possible to actually manufacture a wedge-shaped vibration-damping structure with a tip that is zero thickness, and it is not possible to completely eliminate vibration reflection, so a member that suppresses the vibration is provided at the tip of the wedge.

[0005] Non-Patent Document 1 discloses that by placing a mass on the tip of the wedge via an elastic body, the dynamic damper effect causes the mass to vibrate instead of the tip of the wedge, thereby suppressing vibration of the tip of the wedge. [Prior art documents] [Non-patent literature]

[0006] [Non-Patent Document 1] VIBRATION CONTROL OF VARIABLE THICKNESS PLATES WITH EMBEDDED ACOUSTIC BLACK HOLES AND DYNAMIC VIBRATION ABSORBERS X. Jia, ASME 2015 Noise Control and Acoustics Division Conference at Inter-Noise 2015 Summary of the Invention [Problem to be solved by the invention]

[0007] However, the vibration-damping structure described in Non-Patent Document 1 suppresses vibration by the elastic body expanding and contracting in the amplitude direction of the vibration-damping structure, and it is difficult to suppress low-frequency vibration.

[0008] In other words, since the dynamic damper suppresses the resonance phenomenon around the natural frequency of the wedge, the resonant frequency (f) of the dynamic damper must be reduced in order to suppress low-frequency vibrations, and the resonant frequency (f) of the dynamic damper is given by the following equation (A).

[0009]

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[0010] Therefore, if the elastic body of the dynamic damper is not changed, the spring constant (k) does not change, so the mass (m) of the mass body placed on the elastic body must be increased, which increases the mass of the vibration control structure.

[0011] Furthermore, if the mass (m) of the mass body is to be kept constant, significant design changes are unavoidable in order to reduce the spring constant (k).

[0012] In other words, the spring constant (k) of an elastic body that expands and contracts is determined by the shape of the elastic body and the elastic coefficient of the material that makes up the elastic body, and since the elastic coefficient is a physical property specific to the material, the spring constant of the elastic body cannot be changed without changing the shape and material of the elastic body.

[0013] The present invention was made in consideration of the problems associated with the prior art, and its purpose is to provide a vibration-damping structure that can easily suppress lower-frequency vibrations without increasing the mass of the vibration-damping structure. [Means for solving the problem]

[0014] As a result of extensive research into achieving the above-mentioned object, the inventor discovered that by making the elastic member attached to the tip of the wedge-shaped portion protrude beyond the free end of the wedge-shaped portion, the spring constant of the elastic member can be changed by the length of protrusion, and low-frequency vibrations can be easily suppressed, thereby completing the present invention.

[0015] That is, the vibration damping structure of the present invention includes a vibration damping member having a wedge-shaped portion and an elastic member. The thickness of the wedge-shaped portion from the side where the vibration is incident to the free end where the vibration propagates varies so as to satisfy the following formula (1): The elastic member is a plate-like member having a Young's modulus smaller than that of the vibration-damping member, and is characterized in that a part of the elastic member is fixed so as to protrude beyond the free end of the wedge-shaped portion. h(x) = ε x n + h0...Equation (1) However, in formula (1), x: Distance from the free end (mm) h(x): Thickness at distance x from the free end (mm) h0: Thickness of free end (mm) ε: positive constant n: real number greater than or equal to 1 [Effects of the Invention]

[0016] According to the present invention, an elastic member is provided that protrudes from the free end of the wedge-shaped portion, so that the spring constant of the elastic member can be changed depending on the protruding length of the elastic member, and a vibration-damping structure can be provided that can easily suppress low-frequency vibrations. [Brief explanation of the drawings]

[0017] [Figure 1] FIG. 1 is a diagram illustrating the action of an acoustic black hole. [Figure 2] 1A and 1B are diagrams illustrating a state in which an elastic member of a vibration damping structure of the present invention suppresses displacement of a free end of a vibration damping member. [Figure 3] FIG. 2 is a perspective view of a vibration-damping member, an elastic member, and a metal plate used in the examples and comparative examples. [Figure 4] 10 is a graph showing the results of measuring the inertance of a vibration damping structure of a comparative example. [Figure 5] 10 is a graph showing the results of measuring the inertance of the vibration damping structure of the example. DETAILED DESCRIPTION OF THE INVENTION

[0018] The vibration damping structure of the present invention will now be described in detail. The vibration damping structure of the present invention includes a vibration damping member and an elastic member. The vibration-damping member suppresses vibrations by utilizing the acoustic black hole effect, and has a wedge-shaped portion whose thickness gradually decreases, and a thick portion whose thickness is constant as needed. The elastic member is a plate-like member having a smaller Young's modulus than the vibration-damping member, and a part of the elastic member is fixed to protrude beyond the free end of the wedge-shaped portion.

[0019] The thickness of the wedge-shaped portion of the vibration-damping member from the side where vibration is incident to the free end where vibration propagates varies so as to satisfy the following formula (1). h(x) = ε x n + h0...Equation (1) However, in formula (1), x: Distance from the free end (mm) h(x): Thickness at distance x from the free end (mm) h0: Thickness of free end (mm) ε: positive constant n: real number greater than or equal to 1

[0020] Vibrations (waves) incident on a wedge-shaped portion that changes in accordance with the above formula (1) propagate toward the free end of the wedge-shaped portion, as shown in Figure 1. As the wave approaches the free end, its amplitude increases and its propagation speed slows, so the vibrations can be suppressed by the acoustic black hole effect.

[0021] However, since the lower threshold frequency that can be attenuated by an acoustic black hole is inversely proportional to the square of the length of the wedge shape and proportional to the thickness of the thicker side of the plate, in order to attenuate low-frequency vibrations, it is necessary to extend the length of the wedge-shaped portion, i.e., to increase the installation space for the acoustic black hole.Therefore, the frequencies that can be attenuated by vibration-damping materials are mainly in the high-frequency range of 2000 to 3000 Hz or higher, and it is not easy to attenuate low-frequency vibrations.

[0022] In the vibration damping structure of the present invention, a plate-shaped elastic member is provided at the tip of the free end of the wedge-shaped portion so as to protrude beyond the free end.

[0023] As shown in Figure 2, the plate-shaped elastic member protruding from this free end bends and vibrates when vibrations are applied, and the part of the elastic member protruding from the free end corresponds to the mass of the dynamic damper, so the elastic member acts as a dynamic damper to suppress displacement of the free end of the wedge-shaped portion.

[0024] Here, the spring constant (k2) resulting from the bending stress when the elastic member is bent and deformed is given by the following formula (2). Spring constant (k2) derived from bending stress = 3EI / L 3 ...Equation (2) In the formula (2), E represents the longitudinal elastic modulus (Young's modulus), I represents the second moment of area, and L represents the protrusion length.

[0025] Furthermore, the resonant frequency (f) of the dynamic damper is determined by the spring constant (in the case of the present invention, the spring constant (k2) derived from bending stress) and the mass (m), as described above.

[0026] Therefore, since the vibration control structure of the present invention bends the elastic member rather than expanding and contracting in the amplitude direction of the incident vibration, the spring constant of the dynamic damper is a spring constant (k2) derived from bending stress, and the spring constant (k2) of the elastic member can be changed by the protruding length (L) from the free end.

[0027] In this way, the vibration control structure of the present invention can shift the resonant frequency (f) toward lower frequencies by increasing the protruding length (L) from the free end, and there is no need to change the longitudinal elastic modulus (E) or second moment of area (I) of the elastic member, in other words, the material or shape of the elastic member, so low-frequency vibrations can be suppressed without increasing the mass.

[0028] The protruding length of the elastic member depends on the Young's modulus of the elastic member, but in the case of a flat plate with a constant thickness, low-frequency vibrations can be further suppressed by having at least 1 / 3 of its length protruding beyond the free end.

[0029] There are no particular restrictions on the elastic member as long as it has a Young's modulus smaller than that of the vibration-damping member. -3 ~3×10 -3 (GPa), and examples of such materials include rubber-based materials.

[0030] The vibration-damping member can be made of metal or resin, and the length of the wedge-shaped portion of the vibration-damping member, i.e., the length from the point where the thickness of the vibration-damping member begins to decrease to the free end, should be sufficient to concentrate input vibrations at the tip, and although it depends on the thickness of the thick portion of the vibration-damping member, it is preferably 20 to 2000 mm. The thickness of the thick portion of the vibration-damping member is preferably 0.5 to 50 mm.

[0031] The wedge-shaped portion of the vibration-damping member may have a thickness that gradually decreases toward the free end while satisfying the above formula (1), and both of the one and the other main surfaces may be formed as flat surfaces or curved surfaces, or one main surface may be formed as a flat surface and the other main surface may be formed as a curved surface.

[0032] The overall shape of the vibration-damping member is not particularly limited as long as it has a wedge-shaped portion, and examples thereof include polygonal shapes such as squares and rectangles, as well as circles and ellipses.

[0033] Methods for fixing the elastic member to the vibration-damping member include overlapping the elastic member with the free end tip of the vibration-damping member and adhering them with an adhesive or fastening them with bolts and nuts.

[0034] The vibration damping structure of the present invention can prevent vibrations from tires, engines, etc., and therefore can be suitably used in automobile parts, such as automobile suspensions and automobile panels. [Example]

[0035] The present invention will be described in more detail below with reference to examples, but the present invention is not limited to the following examples.

[0036] [Comparative Example 1] A flat plate of natural rubber (Young's modulus: 2 × 10) with a constant thickness was attached to the free end of the vibration damping member made of hot-rolled steel plate (SS400, Young's modulus: 206 GPa) with the shape shown in Figure 3. -3 The vibration-damping structure was fabricated by laminating the entire damping member with an elastic member made of 1000 GPa (1000 MPa) without allowing it to protrude beyond the free end of the damping member.

[0037] The dimensions of the vibration damping member are length (L V )300mm, width (W V ) 50 mm, and the length of the wedge-shaped part (L V ') is 200 mm, and the thickness of the free end (T V ) is 0.2 mm, and the thickness of the thick part (TV ') is 5 mm, and the change in thickness of the wedge-shaped part, h(x), is (4.8 / 190)x 2 +0.2. The dimensions of the elastic member are length (L E )30mm, width (W E )50mm Thickness(T E ) is 1.2 mm.

[0038] Comparative Example 2 On the elastic member of Comparative Example 1, a length (L M )20mm, width (W M ) 50mm, thickness (T M ) A 0.58 mm metal plate (cold-rolled steel plate, SPCC, Young's modulus: 206 GPa) was bonded to the substrate to create a vibration-damping structure.

[0039] [Example 1] A vibration damping structure was produced in the same manner as in Comparative Example 1, except that the elastic member was attached so as to protrude 5 mm from the free end of the vibration damping member.

[0040] [Example 2] A vibration damping structure was produced in the same manner as in Comparative Example 1, except that the elastic member was attached so as to protrude 10 mm from the free end of the vibration damping member.

[0041] [Example 3] A vibration damping structure was fabricated in the same manner as in Comparative Example 1, except that the elastic member was attached so as to protrude 20 mm from the free end of the vibration damping member.

[0042] <Evaluation> The vibration-damping structures of the above comparative example and example were freely supported by hanging them from a support column with a rubber cord, with the tip of the wedge-shaped portion facing downward, and an acceleration sensor (356A01 manufactured by PCB Co.) was attached 30 mm from the top end of the vibration-damping structure and 10 mm from the side end.

[0043] The area near the acceleration sensor was struck with an impulse hammer (PCB 086C03 (hard tip)), and the force of the impulse hammer (input) and the acceleration of the acceleration sensor (response) were input into an FFT analyzer (Siemens SCADAS III) and subjected to a Fourier transform to obtain the acceleration / force transfer function (inertance). The strike was repeated five times and the inertance was measured by averaging the results. The results of Comparative Examples 1 and 2 are shown in Fig. 4 together with Example 3. The results of Examples 1 to 3 are shown in Fig. 5.

[0044] 4, it can be seen that Example 3 is able to attenuate vibrations over the entire frequency range of 1 to 1000 Hz compared to Comparative Examples 1 and 2. In particular, it was confirmed that Example 3 has a greater damping effect than Comparative Example 2 in the low frequency range around 100 Hz.

[0045] From the graph in Figure 5, it can be seen that in the low frequency range around 100 Hz, Example 3 with a protrusion length of 20 mm has the greatest damping effect, and in the frequency range around 800 Hz, Example 1 with a protrusion length of 5 mm has the greatest damping effect, so it can be seen that the frequency range to be attenuated can be changed by adjusting the protrusion length. [Explanation of symbols]

[0046] 1. Vibration-damping member 11 Wedge-shaped section 12 Free end 13 Thick wall part 2 Elastic member 21 Projection length

Claims

1. A vibration-damping structure including a vibration-damping member having a wedge-shaped portion and an elastic member, The thickness of the wedge-shaped portion from the vibration incident side to the vibration propagating free end varies so as to satisfy the following formula (1): The vibration-damping structure is characterized in that the elastic member is a plate-shaped member having a Young's modulus smaller than that of the vibration-damping member, and a part of the elastic member is fixed so as to protrude beyond the free end of the wedge-shaped portion. h(x) = ε・x ｎ + h ０ ... Equation (1) However, in formula (1), x: Distance from the free end (mm) h(x): Thickness at distance x from the free end (mm) h ０ : Thickness of free end (mm) ε: positive constant n: real number greater than or equal to 1

2. 2. The vibration damping structure according to claim 1, wherein one-third or more of the length of said elastic member protrudes beyond said free end.

3. An automobile part comprising the vibration damping structure according to claim 1 or 2.

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