CN101499273A - Sound absorbing structure and vehicle component having sound absorbing properties - Google Patents

Abstract

The invention provides a sound absorbing structure and a vehicle part with the same. The sound absorbing structure is constituted of a housing having a hollow portion and an opening and a vibration member composed of a board or diaphragm. The vibration member is a square-shaped material having elasticity composed of a synthetic resin and is bonded to the opening of the housing, thus forming an air layer closed inside the sound absorbing structure by the housing and the vibration member. In the sound absorbing structure, when the lateral/longitudinal dimensions of the air layer and characteristics of the vibration member (e.g. a Young's modulus, thickness, and Poisson's ratio) are set such that the fundamental frequency of a vibration occurring in a bending system falls within 5% and 65% of the resonance frequency of a spring-mass system, a vibration mode having a large amplitude occurs in a frequency band lower than the resonance frequency of the spring-mass system, this improving the sound absorption coefficient.

Description

Sound absorption structure and vehicle part with sound absorption characteristics

Technical field

The present invention relates to a kind of sound absorption structure that is applicable to the sound chamber, relate in particular to vehicle part with sound absorption characteristics.

It is the right of priority of the Japanese patent application of 2008-22558,2008-55367,2008-69794,2008-104965,2008-69795,2008-111481,2008-223442,2008-221316,2008-219129 that the application requires application number, and its content all is incorporated into this by reference.

Background technology

Traditionally, various sound absorption structures in as patent documentation 1 and non-patent literature 1, have been developed and disclose.

Patent documentation 1: the open 2006-11412 of Japanese unexamined patent

Non-patent literature 1: by Sho Kimura, Shokokusha Kabushiki Kaisha at " the Architectural Acoustics and Noise PreventionPlans " the 150th page that show on February 20th, 1981

Patent documentation 1 has been lectured a kind of by tabular or membranaceous vibrating mass and be in the sound absorption structure (hereinafter referred to as plate/film vibration sound absorption structure) that air layer in this vibrating mass rear space absorbs sound.In this plate/film vibration sound absorption structure, spring mass system is made of the quality of vibrating mass and the spring ingredient of air layer.Spring mass system has resonance frequency f[Hz] (hertz), according to equation (1), use atmospheric density ρ ₀[kg/m ³] (every cubic metre of kilogram), velocity of sound c ₀Density p [the kg/m of [m/s] (metre per second (m/s)), vibrating mass ³], the thickness t [m] (rice) of vibrating mass and the thickness L[m of air layer] represent resonance frequency.

$f=\frac{1}{2\mathrm{\&pi;}}{\left\{\frac{{\mathrm{\&rho;}}_0{\mathrm{<figure-callout\; id="0"\; label="c"\; filenames="A200910001987E00361.png,A200910001987E00411.png"\; state="\{\{state\}\}">c</figure-callout>}}_0^2}{\mathrm{\&rho;tL}}\right\}}^{1/2}---\left(1\right)$

When thereby the vibrating mass of plate/film vibration sound absorption structure has elasticity and causes elastic vibration, because this elastic vibration and additionally introduced the characteristic of bending system.Non-patent literature 1 has been lectured a kind of sound absorption structure based on architectural acoustics, wherein uses the length a[m on first limit of rectangle vibrating mass according to equation (2)], the length b[m on second limit], the yang type modulus E[N/m of this vibrating mass ²] the Poisson ratio σ [-] (dimensionless) of (every square metre of newton), this vibrating mass and integer p and q calculate the resonance frequency of plate/film vibration sound absorption structure, and resonance frequency is used for acoustic design.

$f=\frac{1}{2\mathrm{\&pi;}}{\{\frac{{\mathrm{\&rho;}}_0{\mathrm{<figure-callout\; id="0"\; label="c"\; filenames="A200910001987E00361.png,A200910001987E00411.png"\; state="\{\{state\}\}">c</figure-callout>}}_0^{21}}{\mathrm{\&rho;tL}}+{[{\left(\frac{p}{a}\right)}^2+{\left(\frac{q}{b}\right)}^2]}^2\left[\frac{{\mathrm{\&pi;}}^4{\mathrm{Et}}^3}{12\mathrm{\&rho;t}(1-{\mathrm{\&sigma;}}^2)}\right]\}}^{1/2}---\left(2\right)$

In equation (2), with the item of spring mass system

Item (follow closely spring mass system item after) addition with bending system; Therefore, resonance frequency becomes and is higher than the resonance frequency of spring mass system, thus the feasible reduction sound absorption crest frequency that is difficult to.

Relation between the resonance frequency of the resonance frequency of spring mass system and the bending system that obtained by the elastic vibration that elasticity caused of plate is not also fully answered; Therefore, in this plate/film vibration sound absorption structure, in low frequency, can not realize the high sound absorptivity.

Summary of the invention

An object of the present invention is to provide a kind of sound absorption structure that effectively absorbs sound by the sound absorption crest frequency in reduction plate/film vibration sound absorption structure.

In one embodiment of the invention, sound absorption structure is by the hollow shell with opening and contain plate or the vibration of membrane parts constitute, its split shed is sealed by vibrating mass, and the crest frequency that wherein absorbs sound is lower than the resonance frequency of spring mass system, described spring mass system is made of the spring ingredient of the air layer in the hollow space of the quality of vibrating mass and shell, and the fundamental frequency of the elastic vibration of described sound absorption crest frequency and vibrating mass and the spring of the air layer in the hollow space of shell are formed the generation of part correlation ground.

Preferably, the fundamental frequency of the elastic vibration of vibrating mass be in spring mass system resonance frequency 5% to 65% between scope in, described spring mass system is made of the spring ingredient of the hollow space of the quality of vibrating mass and shell.Vibrating mass can be fixed on the shell and by shell and support.

Therein the part of the vibrating mass that will place with housing contacts be fixed on the appropriate location and wherein the hollow space of shell have rectangular shape and opening has in the structure of foursquare shape, preferably use the length a[m on foursquare first limit], the yang type modulus E[N/m of vibrating mass ²], the thickness t [m] of vibrating mass, the Poisson ratio σ of vibrating mass and the thickness L[m of hollow space] satisfy inequality (3).

$3<{\left(\frac{1}{a}\right)}^4\frac{{\mathrm{Et}}^{3}L}{1-{\mathrm{\&sigma;}}^2}<550---\left(3\right)$

The hollow space of shell has rectangular shape and opening has in the structure of rectangular shape therein, preferably uses the length a[m on rectangular first limit], with rectangle in length be a[m] the length b[m on vertical second limit, first limit], the yang type modulus E[N/m of vibrating mass ²], the thickness t [m] of vibrating mass, the Poisson ratio σ of vibrating mass and the thickness L[m of hollow space] satisfy inequality (4).

$12<{[{\left(\frac{1}{a}\right)}^2+{\left(\frac{1}{b}\right)}^2]}^2\left[\frac{{\mathrm{Et}}^{3}L}{(1-{\mathrm{\&sigma;}}^2)}\right]<2100---\left(4\right)$

The hollow space of shell has cylindrical shape and opening and has in the structure of circular shape therein, preferably uses the radius R [m] of described opening, the yang type modulus E[N/m of vibrating mass ²], the thickness t [m] of vibrating mass, the Poisson ratio σ of vibrating mass and the thickness L[m of hollow space] satisfy inequality (5).

$40<{\left[{\left(\frac{1}{R}\right)}^2\right]}^2\frac{{\mathrm{Et}}^{3}L}{(1-{\mathrm{\&sigma;}}^2)}<6850---\left(5\right)$

In this connection, can support vibrating mass simply by shell.

Vibrating mass is supported by shell simply so that its displacement is restricted therein, and wherein the hollow space of shell has rectangular shape and opening has in the structure of foursquare shape, preferably uses the length a[m on foursquare first limit], the yang type modulus E[N/m of vibrating mass ²], the thickness t [m] of vibrating mass, the Poisson ratio σ of vibrating mass and the thickness L[m of hollow space] satisfy inequality (6).

$10<{\left(\frac{1}{a}\right)}^4\frac{{\mathrm{Et}}^{3}L}{(1-{\mathrm{\&sigma;}}^2)}<1820---\left(6\right)$

The hollow space of shell has rectangular shape and opening has in the structure of rectangular shape therein, preferably uses the length a[m on rectangular first limit], with rectangle in length be the length b[m on the second vertical limit of this edge of a], the yang type modulus E[N/m of vibrating mass ²], the thickness t [m] of vibrating mass, the Poisson ratio σ of vibrating mass and the thickness L[m of hollow space] satisfy inequality (7).

$40<{[{\left(\frac{1}{a}\right)}^2+{\left(\frac{1}{b}\right)}^2]}^2\left[\frac{{\mathrm{Et}}^{3}L}{1-{\mathrm{\&sigma;}}^2}\right]<7300---\left(7\right)$

The hollow space of shell has cylindrical shape and opening and has in the structure of circular shape therein, preferably uses the radius R [m] of described opening, the yang type modulus E[N/m of vibrating mass ²], the thickness t [m] of vibrating mass, the Poisson ratio σ of vibrating mass and the thickness L[m of hollow space] satisfy inequality (8).

$161<{\left[{\left(\frac{1}{R}\right)}^2\right]}^2\frac{{\mathrm{Et}}^{3}L}{(1-{\mathrm{\&sigma;}}^2)}<27700---\left(8\right)$

Description of drawings

Fig. 1 is the skeleton view that illustrates according to the outward appearance of the sound absorption structure of first embodiment of the invention.

Fig. 2 is the decomposition diagram of sound absorption structure.

Fig. 3 be sound absorption structure that Fig. 1 is shown with and the various suctions that separated by dividing plate of air layer

The planimetric map of acoustic form.

Fig. 4 illustrates its air layer is divided into two-part sound absorption structure by dividing plate exploded perspective

Figure.

Fig. 5 illustrates its air layer is divided into tetrameric sound absorption structure by dividing plate exploded perspective

Figure.

Fig. 6 illustrates the emulation knot of sound absorption structure about the relation between frequency and the acoustical absorption coefficient

The curve map of fruit.

Fig. 7 is the block scheme that the designing apparatus that is used to design sound absorption structure is shown.

Fig. 8 is the process flow diagram that the design process of sound absorption structure is shown.

Fig. 9 be illustrated according to the use of second embodiment of the invention sound absorber vehicle outside

The skeleton view of seeing.

Figure 10 is the side view that vehicle chassis is shown.

Figure 11 is the amplification sectional view of position Pa among Figure 10.

Figure 12 is the decomposition view about Figure 11.

Figure 13 is the skeleton view of outward appearance that the vehicle of sound absorber has been shown according to the use of third embodiment of the invention.

Figure 14 is illustrated on the roof panel for vehicle curve map that the noise on back seat when sound absorber is installed reduces effect.

Figure 15 be according to the use of fourth embodiment of the invention the expansion synoptic diagram of sun visor of sound absorber.

Figure 16 is the sectional view along Figure 15 center line A-A intercepting.

Figure 17 is the sectional view that is installed in the sound absorber in the vehicle rear pillar that illustrates according to fifth embodiment of the invention.

Figure 18 is the sectional view that the vibration of sound absorber shown in Figure 17 is shown.

Figure 19 is the sectional view that illustrates according to the sound absorber in the door that is installed in vehicle of sixth embodiment of the invention.

Figure 20 is the sectional view that the modified example of sound absorber shown in Figure 19 is shown.

Figure 21 is the part cutting planes figure that is installed in the sound absorber in the vehicle floor that illustrates according to seventh embodiment of the invention.

Figure 22 is the diagrammatic sketch that is used to illustrate the sound absorption principle of the sound absorber that comprises a plurality of pipes.

Figure 23 A is the skeleton view that the modified example of the 7th embodiment is shown.

Figure 23 B is the diagrammatic sketch that is illustrated in the side hurdle of the base plate of seeing on the directions X of Figure 23 A.

Figure 24 is the skeleton view of outward appearance that the Vehicular instrument panel of sound absorber has been shown according to the use of eighth embodiment of the invention.

Figure 25 is that it illustrates the inner structure of the instrument panel of having arranged a plurality of sound absorbers along the sectional view of the intercepting of the line X-X among Figure 24.

Figure 26 is the diagrammatic sketch of seeing on the I direction in Figure 25, and it illustrates the layout of a plurality of sound absorbers.

Figure 27 is the skeleton view of outward appearance that the instrument panel of sound absorber has been shown according to the use of the modified example of the 8th embodiment.

Figure 28 is that it illustrates the layout according to a plurality of sound absorbers of modified example along the sectional view of the intercepting of the line Y-Y among Figure 27.

Figure 29 A illustrates the sectional view that wherein is installed on the example of instrument panel inside according to the panel vibration sound absorption structure of ninth embodiment of the invention.

Figure 29 B is the planimetric map of the upside of the instrument panel shown in Figure 29 A.

Figure 29 C is the planimetric map that the example that a plurality of sound absorbers that wherein will form the panel vibration sound absorption structure of being installed instrument panel inside and the left and right directions of vehicle arrange with paralleling is shown.

Figure 29 D illustrates the sectional view that panel vibration sound absorption structure wherein is installed on the example in the pallet (tray) under the glass behind the vehicle.

Figure 29 E illustrates the sectional view of example that panel vibration sound absorption structure wherein is installed on the floor below of vehicle.

Figure 30 A illustrates wherein by each to have arranged the sectional view of example that panel vibration sound absorption structure that a plurality of shells of a plurality of sound absorbers form is installed on the front stall inside of vehicle.

Figure 30 B illustrates wherein by each to have arranged the sectional view of example that panel vibration sound absorption structure that a plurality of shells of a plurality of sound absorbers form is installed on the back seat inside of vehicle.

Figure 31 A is the sectional view that illustrates according to the panel vibration sound absorption structure of first modified example of the 9th embodiment.

Figure 31 B is the sectional view that illustrates according to the panel vibration sound absorption structure of second modified example of the 9th embodiment.

Figure 31 C is the sectional view that illustrates according to the panel vibration sound absorption structure of the 3rd modified example of the 9th embodiment.

Figure 31 D is the sectional view that illustrates according to the panel vibration sound absorption structure of the 4th modified example of the 9th embodiment.

Figure 31 E is the sectional view that illustrates according to the panel vibration sound absorption structure of the 5th modified example of the 9th embodiment.

Embodiment

1. first embodiment

(A) sound absorption structure

Arrive the sound absorption structure of Fig. 6 description with reference to Fig. 1 according to first embodiment of the invention.

Fig. 1 is the outside drawing of sound absorption structure 1-11; Fig. 2 is the decomposition diagram of the essential part of sound absorption structure 1-11.For the formation of present embodiment is described in understandable mode, the size of sound absorption structure 1-11 does not have accurately realistic size.

Sound absorption structure 1-11 is made of shell 10 and vibrating mass 20.The shell of being made by synthetic resin 10 fashions into the square column of hollow, one end opening and other end sealing, and wherein this shell is made of to 12D the bottom 11 and the sidewall 12A that form at the bottom of the shell.

Vibrating mass 20 is by rubber-like synthetic resin being fashioned into the tabular square part that produces, and it is bonded to the opening part of shell 10.Vibrating mass 20 is by bonding and be fixed to the opening part of shell 10, is sealed in the air layer of the inside rear of vibrating mass 20 (or) of this sound absorption structure 1-11 with formation.In the present embodiment, the material of vibrating mass 20 is a synthetic resin, but is not limited thereto.Can adopt such as paper, metal and fiberboard and so on and have elasticity and cause other material of elastic vibration.Vibrating mass 20 not only can be molded into tabular, also can be molded into membranaceous.Make its distortion by applying power to vibrating mass 20, it can recover then, thereby owing to elasticity is vibrated.The shape of the less two-dimensional expansion of its thickness is compared in tabular expression with the rectangular shape of three-dimensional.Membranaceous (for example film shape and sheet) further reduces with the tabular thickness of comparing, and represents that this shape can recover owing to tension force.Compare with shell 10, vibrating mass 20 has low relatively rigidity (promptly low yang type modulus, little thickness and little inferior section moment (secondary sectional moment)) or relative low mechanical impedance, and it is expressed as " 8 * { (bending stiffness) * (superficial density) } ^1/2"; Therefore, vibrating mass 20 has been showed sound absorption function on shell 10.

Among the sound absorption structure 1-11 of the essential structure on have, will use the dividing plate 30 that forms with shell 10 identical materials to be arranged in the air layer air layer is separated into a plurality of parts (following space with each separation is called a chamber (cell)).

Fig. 3 illustrates the sound absorption structure 1-11 that has removed vibrating mass 20, and sound absorption structure 1-12 to 1-15,1-22 to 1-25,1-33 is to 1-35,1-44 to 1-45 with 1-55, their essential structure is identical with the essential structure of sound absorption structure 1-11, their air layer is separated by dividing plate 30, and has removed vibrating mass 20 from them.

In 1-15, form dividing plate 30 at each sound absorption structure 1-12 with the shape of rectangular slab.In sound absorption structure 1-12 as shown in Figure 4, the length on the y direction of dividing plate 30 equals the distance between sidewall 12B and the 12D, and the height of dividing plate 30 equals the height that records between the upper end of sidewall 12A-12D and the bottom 11.

Each sound absorption structure 1-22 to 1-25,1-33 to 1-35,1-44 in 1-45 and 1-55, be the dividing plate 30 separate air layers of lattice-shaped by integral body.In sound absorption structure 1-22 shown in Figure 5, length on the Y direction of whole dividing plate 30 for lattice-shaped equals the distance between sidewall 12B and the sidewall 12D, length on its directions X equals the distance between sidewall 12A and the 12C, and the height of dividing plate 30 equals the height that records between the upper end of sidewall 12A-12D and the bottom 11.

Each sound absorption structure 1-11 has tabular vibrating mass 20 and at the air layer at vibrating mass 20 rears, has therefore formed plate/film vibration sound absorption structure to 1-15.The end of dividing plate 30 on the Z direction bonded to vibrating mass 20, and the other end bonds to bottom 11.

Thereby the resonance frequency that they take place at the resonance that the resonance of spring mass system can not be independent of bending system is near one another so that the resonance of spring mass system is cooperated with the resonance frequency of definite this sound absorption structure mutually with the resonance of bending system in a kind of plate/film vibration sound absorption structure.When the resonance frequency of spring mass system was independent of the resonance frequency of bending system, these two resonance frequencies may influence each other but work independently of each other.

For studying above-mentioned influence, the inventor uses numerical analysis to come emulation to the resonance frequency of sound absorption structure medi-spring quality system, the resonance frequency and the sound absorption crest frequency of bending system.

Table 1 illustrates the simulation result of sound absorption structure 1-11 to 1-55, and table 2 illustrates the sound absorption structure 1-11 that changed the horizontal and vertical length of the chamber simulation result to 1-55.Here, " a " represents the lateral length of each chamber, " b " represents the longitudinal length of each chamber, L represents the thickness of air layer, fb represents the fundamental frequency of spring mass system, fk represents the fundamental frequency of spring bending system, and fk/fb represents the fundamental frequency fk of bending system and the ratio of the fundamental frequency fb of spring mass system, fp representative sound absorption crest frequency.

Table 1

Sound absorption structure	a	b	L	fb	fk	fk/fb(％)	fp
								1-11	315	315	30	385	15	4	380
1-12	156	315	30	385	42	11	180
								1-22	156	156	30	385	61	16	180
1-13	103	315	30	385	90	23	320
								1-23	103	156	30	385	104	27	220
1-33	103	103	30	385	139	36	280
								1-14	77	315	30	385	160	42	360
1-24	77	156	30	385	171	45	260
								1-34	77	103	30	385	199	52	320
1-44	77	77	30	385	250	65	360
								1-15	61	315	30	385	253	66	400
1-25	61	156	30	385	263	68	420
								1-35	61	103	30	385	286	74	380
1-45	61	77	30	385	328	85	420
								1-55	61	61	30	385	394	102	480

Table 2

Sound absorption structure	a	b	L	fb	fk	fk/fb(％)	fp
								(1)	252	336	30	337	10	3	320
(2)	168	252	30	337	21	6	200
								(3)	126	336	30	337	33	10	160
(4)	126	168	30	337	40	12	100
								(5)	112	126	30	337	58	17	160
(6)	84	336	30	337	73	22	260

In the superincumbent emulation, with the thickness L on the Z direction of air layer (promptly and the surface of vibrating mass 20 bottom 11 staggered relatively and and the back side of bottom 11 vibrating mass 20 staggered relatively between distance) be arranged to 30[mm] (millimeter), and the lateral length " a " of each chamber in the sound absorption structure and longitudinal length " b " be arranged to as shown in Table 1 and Table 2 value.In addition, the density of vibrating mass 20 is ρ=940[kg/m ³], the Poisson ratio of vibrating mass 20 is σ=0.4, the thickness of vibrating mass is t=0.85[mm], the yang type modulus of vibrating mass 20 is E=8.8 * 10 ⁸[N/m ²].In table 1 and table 2, calculate the resonance frequency fb of spring mass system with equation (1).To be right after in the equation (2) at first (ρ of spring mass system ₀c ₀ ²/ ρ tL) second fundamental frequency fk that calculates bending system afterwards.In second of equation (2), integer is arranged to p=1 and q=1 (following will call the resonance frequency of the bending system of using p=1 and q=1 to calculate the fundamental frequency of bending system).Produce sound absorption crest frequency fp by the sound absorption characteristics of each sound absorption structure being carried out numerical simulation.Specifically, come together to have determined to arrange the sound field in the sound pipe of sound absorption structure according to JIS A 1405-2 (being entitled as " the determining-second portion of acoustical absorption coefficient and impedance in acoustics-impedance tube: transfer function method (Acoustics-Determination of sound absorption coefficient andimpedance in impedance tubes-Part2:Transfer-function method) ") and finite element method and boundary element method, with calculation of transfer function, thereby calculate sound absorption characteristics.In 1-55, the density p of air layer thickness L, vibrating mass 20 and the thickness t of vibrating mass 20 all are fixed as identical value, at all sound absorption structure 1-11 so that the resonance frequency fb of spring mass system is fixed as identical value.In each of (6), the thickness t of vibrating mass 20 is fixed as identical value, in chamber size sound absorption structure as shown in table 2 (1) so that the resonance frequency fb of spring mass system is fixed as identical value.

As shown in Table 1 and Table 2, the fundamental frequency fk of bending system is more relatively low than the resonance frequency fb of spring mass system, wherein when the fundamental frequency fk of bending system less than the resonance frequency of spring mass system 5% the time (the sound absorption structure 1-11 in the table 1, with and chamber size be 252[mm in the table 2] * 336[mm] sound absorption structure (1)), the vibration of bending system occur in vibrating mass 20 in the approaching frequency place of resonance frequency fb of spring mass system, thereby the Oscillation Amplitude of vibrating mass 20 is owing to its dispersion behavior (dispersed behavior) reduces, and has therefore reduced acoustical absorption coefficient.Can vibrate independently of one another, so the resonance frequency fb of the spring mass system crest frequency (fb ≈ fp here〉〉 fk) that mainly determines to absorb sound because thereby the fundamental frequency fk of bending system is significantly less than two frequencies of resonance frequency fb of spring mass system.In this case, second the value relevant with the fundamental frequency fk of bending system in the equation (2) becomes enough low, with realize that chamber size increases, the yang type modulus of the flexibility of vibrating mass 20, vibrating mass 20 reduces, the thickness of vibrating mass 20 reduces, air layer thickness reduces and superficial density increases.

As shown in table 1, when the fundamental frequency fk of bending system becomes than 65% when high (sound absorption structure 1-15,1-25,1-35,1-45 and 1-55) of the resonance frequency fb of spring mass system, significantly vibration does not appear having in bending system in than the low frequency band of the resonance frequency fb of spring mass system; Therefore, acoustical absorption coefficient can not increase.In addition, the fundamental frequency fk that the resonance frequency fb of spring mass system must be added to bending system goes up increasing sound absorption crest frequency fp, thus acoustical absorption coefficient can not increase in than the low frequency band of the fundamental frequency fk of the resonance frequency fb of spring mass system and bending system (fb and fk＜fp) here.This expresses the sound absorption characteristics by equation (2) decision, realizes therefore that chamber size reduces, the yang type modulus of the hardness of vibrating mass 20, vibrating mass 20 increases, the thickness of vibrating mass 20 increases, air layer thickness increases and superficial density reduces.

When the fundamental frequency fk of bending system is in 5% the scope between 65% of spring mass system resonance frequency fb (the sound absorption structure 1-12 in the table 1 to 1-14,1-22 to 1-24,1-33 is to 1-34 and 1-44, and the sound absorption structure in the table 2 (2) is to (6)), the fundamental vibration of bending system is cooperated with air layer spring ingredient behind, thereby the violent oscillatory motion in the frequency band between the fundamental frequency fk of resonance frequency fb that excites in spring mass system and bending system, so the acoustical absorption coefficient increase (fb〉fp〉fk).

(the sound absorption structure 1-12 in the table 1,1-13,1-22,1-23 and 1-33 when the fundamental frequency fk of bending system is in 5% the scope between 40% of resonance frequency fb of spring mass system, and the sound absorption structure in the table 2 (2) is to (6)), sound absorption crest frequency fp becomes enough lower than the resonance frequency fb of spring mass system.This sound absorption structure is preferably used for absorption frequency and is lower than 300[Hz] sound because the fundamental frequency fk of bending system is owing to low order elastic vibration pattern becomes enough lower than the resonance frequency fb of spring mass system.

The inventor studied the fundamental frequency fk that makes bending system be in spring mass system resonance frequency fb 5% to 65% between scope in condition, determine to have square and its vibrating mass 20 by bonding and be fixed to any sound absorption structure on dividing plate 30 and the shell 10 thus, must satisfy inequality (9) for its chamber.

$3<{\left(\frac{1}{a}\right)}^4\frac{{\mathrm{Et}}^{3}L}{(1-{\mathrm{\&sigma;}}^2)}<550---\left(9\right)$

Obtain inequality (9) by following value and equation.

By using representative about the length " a " on first limit of the α of the different dimensionless factor of vibration mode, vibrating mass, the yang type modulus E of vibrating mass, the thickness t of vibrating mass, thickness L, the Poisson ratio σ of air layer, the density p of vibrating mass, the density p of air layer ₀With airborne velocity of sound c ₀, provide the fundamental frequency fk of bending system by equation (a), and provide the resonance frequency fb of spring mass system by equation (b).

$\mathrm{fk}=\frac{1}{2\mathrm{\&pi;}}\&CenterDot;\mathrm{\&alpha;}\&CenterDot;\frac{t}{a^2}\sqrt{\frac{E}{(1-{\mathrm{\&sigma;}}^2)\mathrm{\&rho;}}}---\left(a\right)$

$\mathrm{fb}=\frac{1}{2\mathrm{\&pi;}}\sqrt{\frac{{\mathrm{\&rho;}}_0{\mathrm{<figure-callout\; id="0"\; label="c"\; filenames="A200910001987E00361.png,A200910001987E00411.png"\; state="\{\{state\}\}">c</figure-callout>}}_0^2}{\mathrm{\&rho;tL}}}---\left(b\right)$

The fundamental frequency fk that inequality (c) satisfies bending system be in spring mass system resonance frequency fb 5% to 65% between scope in condition, and it is launched into inequality (d).

0.05≤fk/fb≤0.65(c)

0.05×fb≤fk≤0.65×fb(d)

Equation (a) and equation (b) substitution inequality (d) are obtained inequality (e).

$0.05\&times;\sqrt{\frac{{\mathrm{\&rho;}}_0{\mathrm{<figure-callout\; id="0"\; label="c"\; filenames="A200910001987E00361.png,A200910001987E00411.png"\; state="\{\{state\}\}">c</figure-callout>}}_0^2}{\mathrm{\&alpha;}}}\&le;\sqrt{\mathrm{tL}\&CenterDot;\frac{t}{a^2}}\&CenterDot;\sqrt{\frac{E}{(1-{\mathrm{\&sigma;}}^2)}}\&le;0.65\&times;\sqrt{\frac{{\mathrm{\&rho;}}_0{\mathrm{<figure-callout\; id="0"\; label="c"\; filenames="A200910001987E00361.png,A200910001987E00411.png"\; state="\{\{state\}\}">c</figure-callout>}}_0^2}{\mathrm{\&alpha;}}}---\left(e\right)$

In the above, " α " 10.40 (sees " actual vibration computing method (Practical Vibration Calculation Method) " version 6 (author: Yoichi Kobori, publisher: Kougaku-Tosho Kabushiki Kaisha) the 213rd page), wherein use ρ ₀c ₀=414 and c ₀=340 are launched into inequality (e) as lower inequality, thereby obtain inequality (9).

$0.05\&times;\frac{375.2}{10.4}\&le;\frac{1}{a^2}\sqrt{\frac{{\mathrm{Et}}^{3}L}{1-{\mathrm{\&sigma;}}^2}}\&le;0.65\&times;\frac{375.2}{10.4}$

$1.80\&le;\frac{1}{a^2}\sqrt{\frac{{\mathrm{Et}}^{3}L}{1-{\mathrm{\&sigma;}}^2}}\&le;23.45$

$3.24\&le;\frac{1}{a^4}\&CenterDot;\frac{{\mathrm{Et}}^{3}L}{1-{\mathrm{\&sigma;}}^2}\&le;549.9$

$3.0<\frac{1}{a^4}\&CenterDot;\frac{{\mathrm{Et}}^{3}L}{1-{\mathrm{\&sigma;}}^2}\&le;550$

For its chamber is thereby that rectangle and its median septum 30 bond to the sound absorption structure that vibrating mass 20 is fixed on the appropriate location, we by emulation find fundamental frequency fk that inequality (10) satisfies bending system be in spring mass system resonance frequency fb 5% to 65% between scope in condition.

$12<{[{\left(\frac{1}{a}\right)}^2+{\left(\frac{1}{b}\right)}^2]}^2\left[\frac{{\mathrm{Et}}^{3}L}{(1-{\mathrm{\&sigma;}}^2)}\right]<2100---\left(10\right)$

Draw inequality (10) by this way: by using finite element method to analyze vibration and analyzing resonance frequency at simple holding state (wherein supporting vibrating mass simply) and stationary state (wherein vibrating mass being fixed on the appropriate location) subsequently.Here, the resonance frequency of simple holding state is 63.7Hz, and the resonance frequency of stationary state is 120.5Hz.The resonance frequency of stationary state is 1.892 with the ratio of the resonance frequency of simple holding state, and its square obtained 3.580, and the value behind this square is used as corrected value.By inequality (12) both sides are all obtained inequality (10) divided by 3.580.

Inequality (9) and (10) show, with respect to the fundamental frequency fk of bending system wherein be in spring mass system resonance frequency fb 5% to 65% between scope in condition for, such as thickness of chamber size, air layer thickness, vibrating mass 20 and so on about the parameter of the size and dimension of vibrating mass 20 and about the material of vibrating mass 20 and the parameter of characteristic the border is arranged all such as yang type modulus, density and Poisson ratio and so on.That is,, can realize acoustic absorption efficiently by being arranged to satisfy inequality (9) and (10) such as the parameter of thickness of chamber size, air layer thickness, vibrating mass 20 and so on and about the material of vibrating mass 20 and the parameter of characteristic.

Fig. 6 illustrates the curve map that measurement result (being entitled as " measuring method of acoustical absorption coefficient in the reverberation chamber (Method formeasurement of sound absorption coefficients in a reverberation room) " according to JIS A 1409 draws with solid-line curve) according to top inequality and actual acoustical absorption coefficient is provided with the simulation result (drawing with dashed curve) of the sound absorption structure of parameter.

In the superincumbent sound absorption structure, the density of vibrating mass 20 is ρ=940[kg/m ³], the Poisson ratio of vibrating mass 20 is σ=0.4, the thickness of vibrating mass 20 is t=0.85[mm], the yang type modulus of vibrating mass 20 is E=8.8 * 10 ⁸[N/m ²], lateral length is 126[mm], longitudinal length is 112[mm], wherein the resonance frequency fb of spring mass system is 471[Hz], the fundamental frequency fk of bending system is 131[Hz], it is 28% of resonance frequency fb.

Fig. 6 is illustrated in sound absorption peak value in the simulation result of sound absorption structure and the measurement result and appears at the low about 315[Hz of resonance frequency fb (being 471Hz) than spring mass system] locate.This shows that simulation result is suitable.

(B) modification

Can revise the first embodiment of the present invention in every way.

In the sound absorption structure of first embodiment, shell 10 has bottom 11, yet can remove bottom 11 from shell 10, wherein forms opening in the side relative with a side that bonds to vibrating mass 20.In this structure, when the opening with shell 10 was fixed to the wall surface in room, the sidewall 12A by wall surface, shell 10 had formed air layer to 12D and vibrating mass 20, thereby had realized plate/film vibration sound absorption structure.Wall by shell 10, vibrating mass 20 and room needs not to be sealing at the sound absorption structure 1-11 inner air layer that forms, and can have small gap or opening.Generally speaking, need owing to sound absorption function is showed in the vibration of the vibrating mass 20 that supports by shell 10.

In the superincumbent modification, vibrating mass 20 is bonding and be fixed on shell 10 and the dividing plate 30, thus limit its bonding portion displacement (or moving); But be not limited thereto.Vibrating mass 20 can also be revised as a kind of simple holding state, this state restriction is shifted from shell 10 but allows around shell 10 rotations.

The inventor finds that inequality (11) satisfies following condition: wherein in having the sound absorption structure of square chamber, the fundamental frequency fk of the bending system that causes owing to elastic vibration be in spring mass system resonance frequency fb 5% to 65% between scope in.

$10\&le;{\left(\frac{1}{a}\right)}^4\frac{{\mathrm{Et}}^{3}L}{1-{\mathrm{\&sigma;}}^2}\&le;1820---\left(11\right)$

Analyze resonance frequency at simple holding state (wherein supporting vibrating mass simply) and stationary state (wherein vibrating mass being fixed on the appropriate location) subsequently by vibrating also, thereby obtain inequality (11) according to finite element method analysis.Here, the resonance frequency of simple holding state is 88Hz, and the resonance frequency of stationary state is 160Hz.The resonance frequency of stationary state is 1.818 with the ratio of the resonance frequency of simple holding state, and its square obtained 3.306, and this square value is used as corrected value.Inequality (9) both sides can obtain inequality (11) by all be multiply by 3.306.

At its chamber is that the inventor finds that inequality (12) satisfies following condition under the situation of rectangle and its vibrating mass 20 sound absorption structure that is in simple holding state: the fundamental frequency fk of the bending system that causes owing to elastic vibration be in spring mass system resonance frequency fb 5% to 65% between scope in.

$40<{[{\left(\frac{1}{a}\right)}^2+{\left(\frac{1}{b}\right)}^2]}^2\left[\frac{{\mathrm{Et}}^{3}L}{1-{\mathrm{\&sigma;}}^2}\right]<7300---\left(12\right)$

According to as get off to draw inequality (12):

The fundamental frequency fk of bending system is by equation (f) expression, and the resonance frequency fb of spring mass system is represented by equation (b).In equation (f), " a " represents the length on the long limit of chamber, and " b " represents the length of the minor face of chamber.

$\mathrm{fk}=\frac{1}{2\mathrm{\&pi;}}\sqrt{\frac{{(1/a^2+1/b^2)}^2{\mathrm{\&pi;}}^4{\mathrm{<figure-callout\; id="3"\; label="Et"\; filenames="A200910001987E00431.png,A200910001987E00432.png"\; state="\{\{state\}\}">Et</figure-callout>}}^3}{12\mathrm{\&rho;t}(1-{\mathrm{\&sigma;}}^2)}}---\left(f\right)$

Wherein the fundamental frequency fk of bending system be in spring mass system resonance frequency fb 5% to 65% between scope in this condition by inequality (g) expression, inequality (g) is expanded into inequality (h).

0.05≤fk/fb≤0.65 (g)

0.05×fb≤fk≤0.65×fb (h)

Equation (f) and equation (b) substitution inequality (h) are obtained inequality (i), and inequality (i) launches to obtain inequality (12).

43.0≤(1/a ²+1/b ²) ²Et ³L(1-σ ²)≤7238 (i)

∴40.0<(1/a ²+1/b ²) ²Et ³L(1-σ ²)<7300

In the present embodiment, to look from above all be square to shell 10 and vibrating mass 20; Yet they needn't be limited as square, and can change into rectangle or other shape.

Can revise present embodiment like this, make shell 10 have the cylindrical shape of end sealing, " circle " opening that wherein discoid vibrating mass 20 is bonded to shell 10 has the outward appearance of columned sound absorption structure with formation.Therein that discoid vibrating mass 20 is bonding and be fixed in the sound absorption structure of shell 10, for the fundamental frequency fk of the bending system that wherein causes owing to elastic vibration be in spring mass system resonance frequency fb 5% to 65% between scope in this condition, the inventor determines to satisfy inequality (13), and wherein R represents the radius of vibrating mass 20.

$40<{\left[{\left(\frac{1}{R}\right)}^2\right]}^2\left[\frac{{\mathrm{Et}}^{3}L}{1-{\mathrm{\&sigma;}}^2}\right]<6850---\left(13\right)$

According to the following inequality (13) that draws:

The fundamental frequency fk of bending system is by radius R that has used vibrating mass and the dimensionless factor α that depends on vibration mode _DcEquation (j) represent that and the resonance frequency fb of spring mass system is represented by equation (b).

$\mathrm{fk}=\frac{1}{2\mathrm{\&pi;}}\&CenterDot;\frac{{\mathrm{\&alpha;}}_{\mathrm{dc}}t}{R^2\sqrt{\frac{E}{\mathrm{\&rho;}(1-{\mathrm{\&sigma;}}^2)}}}---\left(j\right)$

Wherein the fundamental frequency fk of bending system be in spring mass system resonance frequency fb 5% to 65% between scope in this condition represent by inequality (k).Equation (j) and equation (b) substitution inequality (k) are obtained inequality (1).

0.05≤fk/fb≤0.65 (k)

$\frac{0.05}{{\mathrm{\&alpha;}}_{\mathrm{dc}}\sqrt{{\mathrm{\&rho;}}_0{\mathrm{<figure-callout\; id="0"\; label="c"\; filenames="A200910001987E00361.png,A200910001987E00411.png"\; state="\{\{state\}\}">c</figure-callout>}}_0^2}}\&le;\sqrt{\frac{{\mathrm{Et}}^{3}L}{(1-{\mathrm{\&sigma;}}^2){\mathrm{<figure-callout\; id="2"\; label="R"\; filenames="A200910001987E00411.png"\; state="\{\{state\}\}">R</figure-callout>}}^2}}\&le;\frac{0.65}{{\mathrm{\&alpha;}}_{\mathrm{dc}}\sqrt{{\mathrm{\&rho;}}_0{\mathrm{<figure-callout\; id="0"\; label="c"\; filenames="A200910001987E00361.png,A200910001987E00411.png"\; state="\{\{state\}\}">c</figure-callout>}}_0^2}}---\left(1\right)$

Be fixed under the situation of minimum resonance frequency of circle of appropriate location α on its border _DcBe 2.948 (to see " actual vibration computing method (Practical Vibraction CalculationMethod) " version 6 (author: Yoichi Kobori, publisher: Kougaku-ToshoKabushiki Kaisha) the 208th page), wherein use ρ ₀c ₀=414 and c ₀=340 are launched into inequality (1) as lower inequality, thereby obtain inequality (13).

$6.363\&le;\sqrt{\frac{{\mathrm{Et}}^{3}L}{R^2(1-{\mathrm{\&sigma;}}^2)}}\&le;82.72$

$40.49\&le;\frac{{\mathrm{Et}}^{3}L}{(1-{\mathrm{\&sigma;}}^2){\mathrm{<figure-callout\; id="4"\; label="R"\; filenames="A200910001987E00441.png"\; state="\{\{state\}\}">R</figure-callout>}}^4}\&le;6843$

$40.0\&le;\frac{{\mathrm{Et}}^{3}L}{(1-{\mathrm{\&sigma;}}^2){\mathrm{<figure-callout\; id="4"\; label="R"\; filenames="A200910001987E00441.png"\; state="\{\{state\}\}">R</figure-callout>}}^4}\&le;6850$

Discoid therein vibrating mass 20 is simply supported to limit its displacement by shell 10 but allows in the sound absorption structure of its rotation, the fundamental frequency fk that the inventor determines wherein the bending system that causes owing to elastic vibration be in spring mass system resonance frequency fb 5% to 65% between scope in this condition satisfy inequality (14).

$161<{\left[{\left(\frac{1}{R}\right)}^2\right]}^2\left[\frac{{\mathrm{Et}}^{3}L}{1-{\mathrm{\&sigma;}}^2}\right]<27700---\left(14\right)$

Analyze resonance frequency at simple holding state (wherein supporting vibrating mass simply) and stationary state (wherein vibrating mass being fixed on the appropriate location) subsequently by vibrating also, thereby obtain inequality (14) according to finite element method analysis.Here, the resonance frequency of simple holding state is 91Hz, and the resonance frequency of stationary state is 183Hz.The resonance frequency of stationary state is 2.011 with the ratio of the resonance frequency of simple holding state, and its square obtained 4.044, and this square value is used as corrected value.Inequality (13) both sides can obtain inequality (14) by all be multiply by 4.044.

The sound absorption structure that vibrating mass 20 and air layer have all reduced thickness in the present embodiment can not occupy big space in the sound absorption position; Therefore, can realize the absorption of sound with the space that reduces.In order to realize acoustic absorption with the space that reduces, preferably the thickness of vibrating mass 20 less than the thickness of 30mm and air layer less than 30mm.

The sound absorption structure of present embodiment can be arranged to various types of sound chambers.The sound chamber refers to the listening room, meeting room of for example normal room and buildings, soundproof room, hall, theater, audio frequency apparatus, as the designated space of the various transportation systems of vehicle, aircraft and ship and so on and as the inner/outer space of the shell of loudspeaker and musical instrument and so on acoustical generator.

(C) design of sound absorption structure

Can the equipment of using a computer design above satisfying sound absorption structure 1 by the condition of equation and inequality definition.

Fig. 7 illustrates the block diagram that is used for the designing apparatus 50 that is designed by the sound absorption structure 1 of the condition of equation and inequality definition above satisfying.Designing apparatus 50 is made of CPU52, ROM53, RAM54, storer 55, input block 56 and display 57, and all these parts all are connected to together by bus 51.

Storer 55 has hard disk unit, its stored be used for controlling Design equipment 50 with the OS program that realizes operating system and be used for to above satisfying by designing program that the sound absorption structure 1 of the condition of equation and inequality definition designs.Input block 56 has the input media as keyboard and mouse and so on, is used for handling the necessary parameter of user instruction (for example Poisson ratio of the thickness of vibrating mass 20 and size (for example lateral length and longitudinal length, radius etc.), vibrating mass 20 and the yang type modulus of vibrating mass 20) with the design sound absorption structure from designing apparatus 50 inputs.Display 57 has LCDs, and its demonstration is used for input menu that design sound absorption structure necessary parameter is imported, and show satisfy above by the parameter of the condition of equation and inequality definition.

ROM 53 storing initial program loader (IPL).When designing apparatus 50 provides electric power, CPU52 reads IPL from ROM 53 and begins operation.When CPU 52 begins to operate by IPL, read the OS program and carry out it and realize being used to receiving by function of the instruction of input block 56 inputs, be used for screen at display 57 and show the function of various data and image and be used for storer 55 and the function controlled by the basic function that computer equipment is carried out from storer 55.When CPU 52 execution were designed program, designing apparatus 50 inputs were used to design the function of sound absorption structure 1 with realization about the parameter of sound absorption structure 1.

Fig. 8 is the process flow diagram that the part processing of carrying out the designing apparatus 50 of designing program is shown.

When according to the material of predetermined air layer thickness and predetermined vibrating mass 20 and to come vibrating mass 20 wherein according to the specified size of equation above satisfying and inequality be that foursquare sound absorption structure 1 is when designing, the user of designing apparatus 50 operates input block 56, will and storing (step S1) among the RAM54 into such as the parameter input of the thickness of the yang type modulus of air layer thickness, vibrating mass 20 and vibrating mass 20 and Poisson ratio and so on.Then, the parameter that designing apparatus 50 will be stored among the RAM 54 is applied in the top equation and inequality, with the length (step S2) on first limit of calculating vibrating mass 20, thereby shows the length that calculates on the screen of display 57.

As mentioned above, designing apparatus 50 can easily calculate the size of sound absorption structure 1 when receiving the parameter of user's input.Designing apparatus 50 can inputted vibration parts 20 size, yang type modulus and Poisson recently calculate satisfy above the air layer thickness of equation and inequality.Replacedly, the thickness of the vibrating mass 20 of equation and inequality above size, yang type modulus and Poisson ratio that designing apparatus 50 can inputted vibration parts 20 and air layer thickness calculate and satisfy.

Designing apparatus 50 is carried out calculating with the fundamental frequency of generation elastic vibration and the resonance frequency of spring mass system according to the parameter of input, thereby the result that will calculate is presented on the screen of display 57.For example, can calculate these frequencies according to finite element method and boundary element method by designing program.

2. second embodiment

Fig. 9 has illustrated according to the employing of second embodiment of the invention the four-door sedan 100 of sound absorber SA_1) the skeleton view of outward appearance.In vehicle 100, each all is attached to 101, four doors 102 of car bonnet (or hood) and luggage compartment door 103 on the chassis corresponding to the vehicle structure pedestal in the mode of opening/closing.

Figure 10 is the side view that the chassis 110 of vehicle 100 is shown.Chassis 110 has been equipped with base plate 111, from base plate 111 upwardly extending front pillars 112, center pillar 113, rear pillar 114, top 115 (it is supported by post 112,113 and 114), be used for the luggage compartment dividing plate 120 that inner space with vehicle 100 is separated into the engine dividing plate 116 of compartment 105 and engine room 106 and is used for separating between compartment 105 and luggage space 107.Luggage compartment dividing plate 120 has been equipped with back pkt. pallet (rear package tray) 130.

As shown in figure 10, luggage compartment dividing plate 120 comprises the back support of back seat, so cross sectional curve becomes L shaped.

The prerequisite that following description is separated between compartment 105 and luggage space 107 based on luggage compartment dividing plate 120.

Second embodiment is characterised in that box-like sound absorber SA_1 is attached to the luggage compartment dividing plate 120 on chassis 110.Figure 11 is the sectional view of the position Pa among Figure 10, and Figure 12 is the decomposition section that is used to assemble sound absorber SA_1 and luggage compartment dividing plate 120.Figure 11 and Figure 12 illustrate single sound absorber SA_1; In fact, in luggage compartment dividing plate 120 as shown in Figure 9, installed and had difform a plurality of sound absorber SA_1.In this connected, the shape of sound absorber SA_1 was similar or identical with the shape of the luggage compartment dividing plate 120 that is used for separating between compartment 105 and luggage space 107.

As shown in figure 11, back pkt. pallet 130 is attached to luggage compartment dividing plate 120 to form luggage compartment plate 140.

Back pkt. pallet 130 is by constituting with the core 131 that the fabric with acoustical conductivity forms with wood fibre board.Cover the surface of core 131 with surfacing 135.With the part of sound absorber SA_1 core 131 staggered relatively in formed through hole 132 with rectangular aperture.That is, the through hole 132 of surfacing 135 has formed microphone 136, and the acoustic pressure that produces in its compartment 105 transmits to sound absorber SA_1.The opening shape of through hole 132 need not be confined to rectangle, can change over circle.That is, the opening shape of determining through hole 132 is sent to sound absorber SA_1 with the air in compartment 105.

3. the 3rd embodiment

With reference to Figure 13 and 14 the third embodiment of the present invention is described.In Figure 13, the building block identical with the parts shown in Fig. 9 and 10 specified with identical reference number.

Figure 13 is the skeleton view of outward appearance that the four-door sedan 100 of sound absorber SA_2 has been shown according to the employing of third embodiment of the invention.Each all is attached to car bonnet 101, four doors 102 and luggage compartment door 103 on the chassis 110 corresponding to the pedestal of vehicle structure in the mode of opening/closing.Form the chassis 110 of vehicle 100 as shown in figure 10.Compare with second embodiment that wherein sound absorber SA_1 is attached on the back pkt. pallet 130, the 3rd embodiment is designed to sound absorber SA_2 is attached to top 240.Top 240 is made of top outside plate (corresponding to the top among Figure 10 115) and top inner panel 230.

The 3rd embodiment is characterised in that box-like sound absorber SA_2 is attached to the top 240 of vehicle 100.In Figure 13, sound absorber SA_2 comprises four sound absorber SA_2a and SA_2b with different size altogether.

In top 240, with the top outside plate of inner panel 230 clips in top to formation chassis 110 parts.

In top inner panel 230, the surface coverage of the core 231 that constitutes by wood fibre board the surfacing 238 that constitutes by fabric with acoustical conductivity.Form rectangular through-hole 232A near the core 231 of back seat, wherein the part with through hole 232A surfacing 238 staggered relatively has formed microphone 239A.Sound absorber SA_2 is communicated with compartment 105 via microphone 239A.Microphone 239A needn't be attached to the top 240 approaching with back seat, can also change into to be attached to the top approaching with front stall.Figure 14 is the curve map that the noise that is illustrated in the back seat place reduces effect.

4. the 4th embodiment

The 4th embodiment is characterised in that the sun visor 330 that box-like sound absorber SA_3 is attached to vehicle 100.Figure 15 is the stretch-out view of sun visor 330 on top that is attached to the top 115 of vehicle 100, and Figure 16 is the sectional view along the intercepting of the line A-A among Figure 15.

Sun visor 330 is made of every light (light insulation) part 340 and the L shaped support bar 350 that is used for so that rotatable mode supports every light part 340 tabular.

Constitute by the core 341 that forms with ABC resin (or engineering plastics) with the surfacing 360 that the adhesive-bonded fabric with acoustical conductivity forms every light part 340.Cover core 341 so that each limit of surfacing 360 combines to cover the surface and the back side of core 341 with surfacing 360.

An end that is used for sun visor 330 is attached to top 115 carriage 351 and support bar 350 fuses.In carriage 351, form a pair of screw hole 352.By the precalculated position that carriage 351 is threaded onto top 115 sun visor 330 is fixed to top 115.

In core 341, be formed for the rectangular through-hole 342 of attached sound absorber SA_3.The through hole 342 of surfacing 360 is as microphone 361.

5. the 5th embodiment

The 5th embodiment is characterised in that box-like sound absorber SA_4 is attached to rear pillar 114.In fact, can be attached to rear pillar 114 with having difform a plurality of sound absorber SA_4.

Figure 17 is the cross-sectional view that is attached to the sound absorber SA_4 of rear pillar 114.Rear pillar 114 has been equipped with back outside plate 420 (it forms the part on chassis 110) and back inner panel 430 (it is attached to back outside plate 420).

The plate part 421 that use has the rectangular shape of trapezoid cross section forms back outside plate 420.Pilot hole 422 that back inner panel 430 is installed and the protrusion mounting hole 423 that sound absorber SA_4 is installed in plate part 421, have been formed.Back glass 117 is arranged in an end of outside plate 420 afterwards by the sealing (not shown), door glass 118 is arranged in the other end of back outside plate 420 by the sealing (not shown).

Back inner panel 430 constitutes by the core 431 that forms with acrylic resin with the surfacing 439 that the fabric with acoustical conductivity forms, and wherein covers the surface of core 431 with surfacing 439.

Core 431 is made of circular portion 432 and sloping portion 433 (it extends towards circular portion 432 outsides).In circular portion 432, form a plurality of through holes 434.Rear pillar 114 is communicated with compartment 105 via through hole 434.

Figure 18 illustrates the modification of the 5th embodiment, wherein sound absorber SA_4 is inserted in the rectangular depression 436 of the core 431 of compartment 105 inner openings.Pilot hole 436A is formed on the bottom in depression 436.It is inner and its protrusion inserts among pilot hole 436A that sound absorber SA_4 is installed in depression 436.

Present embodiment is designed so that sound absorber SA_4 is attached to rear pillar 114; But this is not restriction.For example, sound absorber SA_4 can be attached in front pillar 112 or the center pillar 113.

6. the 6th embodiment

The 6th embodiment is characterised in that box-like sound absorber SA_5 is attached in the door 102 of vehicle 100.

The inside of door 102 comprises door internal decoration plate matrix 520, internal material 530, handrail 540 and door pocket 550.Internal material 530 is by the door internal decoration plate matrix 520 that forms with synthetic resin and comprise that the surfacing 535 of the adhesive-bonded fabric with acoustical conductivity constitutes.Cover the surface of door internal decoration plate matrix 520 with surfacing 535.

Figure 19 illustrates sound absorber SA_5 is installed in handrail 540 inside, is communicated with a plurality of through hole 520A in being formed on door internal decoration plate matrix 520.

Figure 20 illustrates a plurality of sound absorber SA_5 is installed in internal material 530 inside, be communicated with a plurality of through hole 520A, and another sound absorber SA_5 is used for door pocket 550.

7. the 7th embodiment

The 7th embodiment is characterised in that the sound absorber SA_6 that will comprise a plurality of absorbing ducts is installed in the base plate 111 of vehicle 100.As shown in figure 21, sound absorber 630 (being sound absorber SA_6) is installed in the depression 600 that is formed in the base plate 111.

By being interconnected and fuse, linearly aligned a plurality of pipes 631 (for example 631-1 is to 631-9) with different length form sound absorber 630.Each pipe 631 all is to be made of and the cross section is circular linear rigid pipe synthetic resin.One end of each pipe 631 all seals with the form of enclosure portion 632, and the other end is opened with the form of opening (as microphone) 633, and wherein inside of each pipe 631 is hollow space 634.The opening 633 of each pipe 631 is via being communicated with compartment 105 in the slit of closing formation in 102 o'clock to the doorstep.

Figure 22 illustrates hollow space and has the adjacent tubes 631-i of different length L1 and L2 and the relation between the 631-j.(L1=λ 1/4 here for wavelength X 1 and λ 2, L2=λ 2/4) is four times sound wave generation standing wave S1 and the S2 of length L 1 and L2, near thereby the vibration that causes repetitive propagation among pipe 631-i and the 631-j is with sound energy consumption, thereby the acoustic absorption realization wavelength X 1 and the λ 2.

Figure 23 A illustrates the modification of the 7th embodiment, wherein manages 631 and is arranged in the side hurdle 601 of base plate 111 so that manage 631 hollow space 634 and extend along the fore-and-aft direction of vehicle 100.Figure 23 B is the synoptic diagram from the side hurdle 601 that the directions X of Figure 23 A is watched.

8. the 8th embodiment

The 8th embodiment is characterised in that sound absorber SA_8 is installed in the instrument panel of being arranged below the front glass 105F in the compartment 105 of vehicle 100 700.

Figure 24 is the skeleton view that the outward appearance of instrument panel 700 is shown.Sound absorber SA_8 is arranged in the space S between instrument panel 700 and the engine dividing plate 116.

Instrument panel 700 has been equipped with the loudspeaker 701 and 702 and warm/cold air outlet 703 of various instrument, audio devices.On the upper surface of instrument panel 700, form the warm air that a plurality of defrosters 704 provide from air-conditioning unit 705 with output.On the position, lower-left of instrument panel 700, arrange glove compartment 707 and with cover plate 708 sealings.

Figure 25 illustrates the inner structure of instrument panel 700 and is the sectional view that intercepts along the line X-X among Figure 24.Air-conditioning unit 705, defrosting air pipe 706 and a plurality of sound absorber SA_8A in the internal space S of instrument panel 700, have been arranged.The internal space S of instrument panel 700 is communicated with compartment 105 via hole H.

Figure 26 is the synoptic diagram of the instrument panel 700 watched with the I direction among Figure 25, and it is illustrated in the layout of the sound absorber SA_8A in the last diagrammatic sketch.A plurality of sound absorber SA_8A are arranged in the wide region zone on the upside of inwall of instrument panel 700.In addition, sound absorber SA_8A is arranged to the other parts of the inwall of defrosting air pipe 706 and instrument panel 700 approaching.

Figure 27 is the outward appearance skeleton view that the instrument panel 700 of sound absorber SA_8A has been shown according to the use of the modification of the 8th embodiment.A loudspeaker SP and two sound absorber SA_8B are arranged in together in each of the right side of upper surface of instrument panel 700 and left side.Figure 28 is that it illustrates the inner structure of instrument panel 700 along the sectional view of the intercepting of the line Y-Y among Figure 27.In each of the right side of the upper surface of instrument panel 700 and left side, all formed depression 730.A loudspeaker SP and two sound absorber SA_8B are arranged in the inside of depression 730 together, cover the opening of depression 730 with net N.Also on the inwall of instrument panel 700, arranged another sound absorber SA_8B.In this structure, sound absorber SA_8B has consumed from the compartment 105 and has passed the acoustic energy of coming and the energy of the engine sound that sends from engine room 106 via engine dividing plate 116, thereby realizes acoustic absorption.

Hereinbefore, will not be arranged in the depression 730 of support speaker SP by sound absorber SA_8B; Therefore, sound absorber SA_8B can be arranged in other space of placing instrument etc.Needn't cover sound absorber SA_8B with net N; Therefore, can rearrange sound absorber SA_8B is communicated with compartment 105 via grid, guard and slice.

9. the 9th embodiment

The 9th embodiment is characterised in that by making up a plurality of sound absorbers and forms three-dimensional sound absorption structure.

Specifically, be included in a plurality of sound absorbers 820 in its shell 810 according to the panel vibration sound absorption structure 800 of the 9th embodiment.

With reference to Figure 29 A the example that present embodiment is attached to each position of vehicle 100 is described to 29E.Figure 29 A is the sectional view that has been equipped with the instrument panel 700 of panel vibration sound absorption structure 800, and Figure 29 B is the top planimetric map of instrument panel 700.

Shown in Figure 29 A and 29B, the shell 810 of panel vibration sound absorption structure 800 is attached to the lower position of instrument panel 700, the boundary that wherein approaches front glass 105F in instrument panel 700 forms the elongated hole 733 that elongates in the vertical, and covers with grid G 1.Shell 810 is crooked in the vertical, and its opening has the identical size of essence with the elongated hole 733 of instrument panel 700.That is, panel vibration sound absorption structure 800 is attached on the lower position of instrument panel 700 so that the elongated hole 733 of the opening of shell 810 and instrument panel 700 is relatively placed.

A plurality of sound absorbers 820 are arranged in make its vibration surface vertical in the shell 810 with the virtual plane of the opening that edge of opening surrounded of shell 810.Specifically, the vibration surface of sound absorber 820 and the fore-and-aft direction of vehicle 100 are arranged abreast that wherein sound absorber 820 is arranged in along in the shell 810 of the elongated hole 733 of the instrument panel 700 on vehicle 100 left and right directions.

By in shell 810 with on the corresponding per unit area of the surface area of sound absorber 820, arranging two or more sound absorbers 820, can realize having the panel vibration sound absorption structure 800 of high acoustic absorption efficient.Preferably the panel vibration sound absorption structure 800 with present embodiment is arranged in the pre-position that acoustic pressure increases easily in the vehicle 100.Owing to sound absorber 820 is arranged in the shell 810 so that vibration surface is crossed over the plane of the opening of shell 810, therefore can suitably changes the arranged direction of sound absorber 820.In Figure 29 C, a plurality of sound absorbers 830 are arranged in the shell 810 of panel vibration sound absorption structure 800 so that the left and right directions of its vibration surface and vehicle 100 is placed abreast.Certainly, it is vertical with the plane of the opening of shell 810 sound absorber 820 and 830 can be placed as the vibration surface that makes them.

Figure 29 D illustrates the example as the shell 811 of panel vibration sound absorption structure 800 of the pallet 117T below the back glass of vehicle 100 wherein.Opening with grid G 2 covering shells 811.A plurality of sound absorbers 840 are arranged in the shell 811 with the noise in the back seat that effectively reduces vehicle 100.

Figure 29 E illustrates the example that wherein shell 812 of panel vibration sound absorption structure 800 is arranged in base plate 111 belows of vehicle 100.Base plate 111 has been equipped with porous metals to realize acoustical conductivity, wherein carpet 111C is attached to the upper surface of base plate 111.The below that shell 812 is attached to base plate 111 makes its opening towards base plate 111.In order to strengthen acoustically effective, felt F is adhered to the bottom of shell 812, and cover the pugging SP that constitutes by rubber, thereby on pugging SP, arrange a plurality of sound absorbers 850.In this structure, can effectively reduce the road rumble that enters from vehicle 100 belows the compartment 105.

Figure 30 A illustrates a plurality of shell 815a, 815b and 815c is installed in panel vibration sound absorption structure 800A among the front stall 100F of vehicle 100.In front stall 100F, form the opening (drawing) of grid with dotted line near the opening part of shell 815a, 815b and 815c.In shell 815a, arrange a plurality of sound absorber 860a; In shell 815b, arrange a plurality of sound absorber 860b; In shell 815c, arrange a plurality of sound absorber 860c.In this structure, can absorb the noise in the compartment 105, and can reduce from the acoustic energy of front stall 100F biography to human body.

Figure 30 B illustrates wherein and will import the panel vibration sound absorption structure 800B that be installed among the back seat 100R example with effective absorption sound such as the sound wave of noise.The unitary construction of the panel vibration sound absorption structure 800B roughly structure with panel vibration sound absorption structure 800A is identical.Form opening 800P in the top in the space that forms at the back of the back support of back seat 100R, wherein the open communication of this space and shell 815b.When sound wave when entering the back of back seat 100R with the approaching opening 800P of back seat 100R, can effectively suppress these sound waves.

Next, the modification of present embodiment will be described in conjunction with Figure 31 A at the layout of the sound absorber 920 in the shell 910 of panel vibration sound absorption structure 900 to 31E.

Figure 31 A illustrates a plurality of sound absorber 920A is arranged among the shell 910A of panel vibration sound absorption structure 900A.Sound absorber 920A has support component 940A, and each support component all has and removed two relative sides and the hexahedral shape of remaining four sides, and wherein each the center with four sides forms a surface perpendicularly.When with four sides in the perpendicular direction in a pair of opposite flank on and with another direction that opposite flank is paralleled on when support component 940A cut, its cross sectional shape is roughly H shape.Because the top shape of support component 940A, so opening is formed on the relative two ends of each side, wherein sound absorber 920A made up to make each opening be connected with each vibrating mass 930A.

Side at shell 910A forms an opening.The vibration surface of vibrating mass 930A is arranged in the virtual plane of the opening that edge of opening surrounded of leap by shell 910A.This feasible quantity that can easily regulate the sound absorber 920A among the shell 910A that is arranged in panel vibration sound absorber structure 900A, thus acoustical absorption coefficient improved.

Can make the inclined position of the sound absorber 920A that arranges at the panel vibration sound absorption structure 900A neutral line shown in Figure 31 A.Figure 31 B illustrates a kind of panel vibration sound absorption structure 900B that is encapsulated among the shell 910B, has wherein in position arranged a plurality of sound absorber 920B and has made it inclination.This feasible total area that can reduce highly can not reduce the vibration surface of sound absorber 920B.Therefore, can realize having panel vibration sound absorption structure 900B than low height and high acoustic absorption coefficient.

Can use a sheet material to form a plurality of vibrating mass.Similar with the panel vibration sound absorption structure 900A shown in Figure 31 A, in the shell 910C of panel vibration sound absorption structure 900C, arranged a plurality of support component 940C, wherein support component 940C links together, simultaneously by opening is sealed in a sheet material bending.This has produced by the opening restriction site of support component 940C and has been used to form vibrating mass 930C to absorb the platy structure of sound.This structure allows to form a plurality of sound absorber 920C that have been equipped with a plurality of vibrating mass 930C with a sheet material; Therefore, can easily produce panel vibration sound absorption structure 900C.

Can provide different shapes for the support component 940A of the sound absorber 920A shown in Figure 31 A.In the panel vibration sound absorption structure 900D shown in Figure 31 D, the bottom that plate-like support member 940D is attached at shell 910D is with the opening towards top.The sheet material of bending is attached at the bottom of end and the shell 910D of support component 940D, thereby forms the vibrating mass 930D that supports by support component 940D.This structure allows to be formed on a plurality of sound absorber 920D that have been equipped with a plurality of vibrating mass 930D in the shell 910D with a sheet material; Therefore, can easily produce panel vibration sound absorption structure 900D.

Because the support component of use sound absorber supports vibrating mass and forms air layer in the one side, therefore needn't form air layer in the peripheral region of support component.Figure 31 E illustrates panel vibration sound absorption structure 900E, wherein with the perpendicular direction in each side of shell 910E and bottom surface on sound absorber 920E is cut.

A pair of opposite flank and support component 940E that Figure 31 E illustrates sound absorber 920E relatively place, and illustrate: in this side the opposite flank, from and perpendicular to the scope between contacted position to, the plane vibrating mass 930E at every side center in support component 940E is excised, and in another side, from and contacted position, described plane in the scope another vibrating mass 930E, support component 940E is excised.The sound absorber 920E and the vibrating mass 930E that have promptly partly excised support component 940E are combined into one, and are fixed on the central authorities of the sidewall of shell 910E.In the panel vibration sound absorption structure 900E of Figure 31 E, sound absorber 920E is made of vibrating mass 930E and support component 940E.

In Figure 31 E, support component 940E is fixed to shell 910E side wall centers forming air layer between vibrating mass 930E and support component 940E, has also formed relatively large air layer in (promptly above the bottom of shell 910E) below vibrating mass 930E and the support component 940E simultaneously.This structure allows easily to regulate the cumulative volume of air layer, thereby easily regulates the frequency band that is absorbed sound.

The vibrating mass shape of the sound absorber in the panel vibration sound absorption structure need not be confined to square, can also be varied to different shape, as polygon, circle and oval.In addition, can control the frequency band of sound absorption by extra formation hole in vibrating mass and support component.

At last, the present invention is not limited to top embodiment and modification, can also carry out other modification within the scope of the invention of claims definition.

Claims (13)

1. sound absorption structure comprises:

Shell with hollow space and opening; With

The vibrating mass that constitutes by plate or film,

Wherein cover the opening of described shell with described vibrating mass, and

The crest frequency that wherein absorbs sound occurs in fundamental frequency when the elastic vibration of described vibrating mass with the spring ingredient of the air layer that forms co-operating the time in the hollow space of described shell, described sound absorption crest frequency is lower than the resonance frequency of spring mass system, and described spring mass system is based on the spring ingredient of the air layer of the hollow space of the quality of described vibrating mass and described shell.

2. sound absorption structure as claimed in claim 1, the fundamental frequency of the elastic vibration of wherein said vibrating mass be in described spring mass system resonance frequency 5% to 65% between scope in, described spring mass system is based on the spring ingredient of the air layer of the hollow space of the quality of described vibrating mass and described shell.

3. sound absorption structure as claimed in claim 2, wherein said vibrating mass is fixed on the described shell.

4. sound absorption structure as claimed in claim 3, thereby the hollow space of wherein said shell has the described opening of rectangular shape is square, and wherein use length " a " rice on described foursquare first limit, yang type modulus " E " newton of described vibrating mass/square metre, the thickness " L " of the hollow space of the Poisson ratio " σ " of thickness " t " rice of described vibrating mass, described vibrating mass and described shell meter sets up as lower inequality

$3<{\left(\frac{1}{a}\right)}^4\frac{{\mathrm{Et}}^{3}L}{1-{\mathrm{\&sigma;}}^2}<550.$

5. sound absorption structure as claimed in claim 3, it is rectangle that thereby the hollow space of wherein said shell has the described opening of rectangular shape, and wherein use described rectangle first limit length " a " rice, with described rectangle in length for length " b " rice on the second vertical limit of first limit of " a " rice, yang type modulus " E " newton of described vibrating mass/square metre, the thickness " L " of the hollow space of the Poisson ratio " σ " of thickness " t " rice of described vibrating mass, described vibrating mass and described shell meter sets up as lower inequality

$12<{[{\left(\frac{1}{a}\right)}^2+{\left(\frac{1}{b}\right)}^2]}^2\left[\frac{{\mathrm{Et}}^{3}L}{1-{\mathrm{\&sigma;}}^2}\right]<2100.$

6. sound absorption structure as claimed in claim 3, thereby the hollow space of wherein said shell has the described opening of cylindrical shape for circular, and wherein use yang type modulus " E " newton of the radius R rice of described opening, described vibrating mass/square metre, the thickness " L " of the hollow space of the Poisson ratio " σ " of thickness " t " rice of described vibrating mass, described vibrating mass and described shell meter sets up as lower inequality

$40<{\left[{\left(\frac{1}{R}\right)}^2\right]}^2\frac{{\mathrm{Et}}^{3}L}{1-{\mathrm{\&sigma;}}^2}<6850.$

7. sound absorption structure as claimed in claim 2, wherein said vibrating mass is supported simply by described shell.

8. sound absorption structure as claimed in claim 7, thereby the hollow space of wherein said shell has the described opening of rectangular shape is square, and wherein use length " a " rice on described foursquare first limit, yang type modulus " E " newton of described vibrating mass/square metre, the thickness " L " of the hollow space of the Poisson ratio " σ " of thickness " t " rice of described vibrating mass, described vibrating mass and described shell meter sets up as lower inequality

$10<{\left(\frac{1}{a}\right)}^4\frac{{\mathrm{Et}}^{3}L}{1-{\mathrm{\&sigma;}}^2}<1820.$

9. sound absorption structure as claimed in claim 7, it is rectangle that thereby the hollow space of wherein said shell has the described opening of rectangular shape, and wherein use described rectangular first limit length " a " rice, with described rectangle in length for length " b " rice on the second vertical limit of first limit of " a " rice, yang type modulus " E " newton of described vibrating mass/square metre, the thickness " L " of the hollow space of the Poisson ratio " σ " of thickness " t " rice of described vibrating mass, described vibrating mass and described shell meter sets up as lower inequality

$40<{[{\left(\frac{1}{a}\right)}^2+{\left(\frac{1}{b}\right)}^2]}^2\left[\frac{{\mathrm{Et}}^{3}L}{1-{\mathrm{\&sigma;}}^2}\right]<7300.$

10. sound absorption structure as claimed in claim 7, thereby the hollow space of wherein said shell has the described opening of cylindrical shape for circular, and wherein use yang type modulus " E " newton of the radius R rice of described opening, described vibrating mass/square metre, the thickness " L " of the hollow space of the Poisson ratio " σ " of thickness " t " rice of described vibrating mass, described vibrating mass and described shell meter sets up as lower inequality

$161<{\left[{\left(\frac{1}{R}\right)}^2\right]}^2\frac{{\mathrm{Et}}^{3}L}{1-{\mathrm{\&sigma;}}^2}<27700.$

11. the designing apparatus at sound absorption structure, described sound absorption structure is formed by the shell with hollow space and opening and by the vibrating mass that plate or film constitute, so that seal the opening of described shell by described vibrating mass,

Described designing apparatus comprises:

Input block is used for importing at least one parameter of described vibrating mass and the thickness of the air layer that forms at the hollow space of described shell; With

The unit is set, is used for coming at least the thickness of the air layer of the hollow space of size, yang type modulus, Poisson ratio and the described shell of described vibrating mass any one is provided with according to the parameter of input.

12. designing apparatus as claimed in claim 11 wherein calculates the fundamental frequency of described vibrating mass by the mode of numerical analysis.

13. a sound chamber has according to any described sound absorption structure in the claim 1 to 10.

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