A sound-absorbing super-structured sound barrier structure design method

Abstract

A method for designing a sound-absorbing metamorphic sound barrier structure, wherein the sound-absorbing metamorphic sound barrier structure includes sound-absorbing units assembled within a space enclosed by a back panel and a front panel. Each sound-absorbing unit includes a back cavity and a bottom perforated panel (i.e., the first layer) disposed within the back cavity, a top perforated facing panel (i.e., the nth layer), and one or more sound-absorbing panels (i.e., the second to the (n-1)th layers) located between the first and nth layers. Each layer has a cavity beneath it. The sound absorption coefficient α of the entire sound-absorbing unit is calculated using the following formula: where n is an integer greater than 2, Z... s,n The sound impedance of the uppermost perforated panel (i.e., the nth layer) of the sound-absorbing unit is ρ0, where ρ0 is the air density and c0 is the speed of sound in air. This invention allows for the theoretical calculation of a suitable structure for designing a sound-absorbing metamorphic sound barrier, eliminating the need for direct model fabrication and significantly reducing design difficulty and cost.

Description

Technical Field

[0001] This invention relates to a design method for a sound-absorbing structure, specifically a design method for a sound-absorbing metastructure sound barrier. Background Technology

[0002] Noise pollution has become a global environmental hazard, posing a severe challenge to public health, quality of life, and even socio-economic development. In recent years, with the deepening of my country's ecological civilization construction and the proposal of the "Beautiful China" goal, the country's attention to noise pollution prevention and control has reached an unprecedented level. Against this backdrop, the public's requirements for acoustic environmental comfort are undergoing a qualitative leap. Modern urban residents not only pursue "clear hearing" but also yearn for "quiet hearing"—a tranquil living space, a low-noise office environment, and a smooth and comfortable transportation cabin have become important benchmarks for measuring quality of life. However, traditional noise control technologies are facing significant bottlenecks: porous sound-absorbing materials are effective in the mid-to-high frequency range, but their effectiveness decreases sharply for low-frequency noise (<1000Hz) with longer wavelengths and strong penetration, and they often require a large amount of space; resonant sound-absorbing structures (such as perforated plates and Helmholtz resonators) can target specific low frequencies, but their effective sound absorption bandwidth is extremely narrow, making it difficult to cope with the complex and ever-changing actual noise spectrum. In engineering applications, the demand for new sound-absorbing solutions that combine lightweight structure, high efficiency at low frequencies, and wide bandwidth has never been more urgent.

[0003] Noise barriers, as structures installed on one or both sides of transportation lines such as high-speed railways, urban rail transit, highways, and viaducts, serve two main purposes. First, they absorb some of the noise from transportation vehicles, reducing noise on the side of the vehicles. Second, they block noise, reducing environmental noise from residents and buildings along the line outside the noise barrier. Their application is widespread and mature.

[0004] Existing sound barriers are mainly divided into two types based on their internal sound-absorbing materials: cotton and aluminum foam.

[0005] Cotton-based sound barriers still have unresolved issues: First, due to the pores in porous fiber sound-absorbing materials, the cotton collapses under the influence of gravity, wind loads, and material aging. The collapsed areas lose their sound-absorbing function, further reducing the overall sound absorption performance of the sound barrier. Second, under the current constraints on sound barrier thickness (generally <200mm), porous fiber sound-absorbing materials have poor sound absorption performance in the low-frequency range below 1000Hz, especially below 500Hz, which cannot meet the requirements for low-frequency noise reduction such as wheel-rail noise.

[0006] Sound barriers made of foamed aluminum generally have a narrow sound absorption band, exhibiting good sound absorption in the 200Hz~1400Hz range. Outside these frequencies, the sound absorption performance drops significantly, resulting in poor overall sound absorption performance. Of course, by thickening the foamed aluminum, such as using approximately 120mm thick foamed aluminum, better sound absorption performance can be achieved, but due to the high density of foamed aluminum (approximately 900~1100 kg / m³), this is problematic. 3 This leads to a significant increase in weight and cost. Summary of the Invention

[0007] To address the aforementioned problems, this invention proposes a structural design method for a sound-absorbing metamorphic sound barrier. The resulting metamorphic sound barrier can effectively broaden the sound absorption frequency band and significantly improve the design speed.

[0008] The technical means adopted by this invention to solve the above problems is as follows: a sound-absorbing metamorphic sound barrier structure design method, wherein the sound-absorbing metamorphic sound barrier structure includes a sound-absorbing unit assembled in a space enclosed by a back panel and a front panel. The sound-absorbing unit includes a back cavity and a perforated plate (i.e., the first layer plate) disposed in the back cavity at the bottom, a perforated protective plate (i.e., the nth layer plate) at the top, and one or more sound-absorbing plates (i.e., the second layer plate to the (n-1)th layer plate) located between the first layer plate and the nth layer plate. Each layer plate has a cavity below it. The sound absorption coefficient α of the entire sound-absorbing unit is calculated according to the following formula: , Where: n is an integer greater than 2, Z s,n Let ρ0 be the sound impedance of the uppermost perforated panel of the sound-absorbing unit, i.e., the nth layer, and let c0 be the air density and the velocity of sound in air. Also: , Where: Z s,n’ Let be the acoustic impedance of the lower surface of the nth layer plate, j be the imaginary unit, κ0 be the wave number of air, and t be the acoustic impedance of the lower surface of the nth layer plate. n Let d be the thickness of the nth layer. n Let Φ be the diameter of the perforation hole in the nth layer plate. n Let be the perforation rate of the nth layer, ω be the angular frequency, η be the dynamic viscosity coefficient of air, J0 be the 0th-order Bessel function, and J2 be the 2nd-order Bessel function; while , Where: Z s,n-1 h is the surface acoustic impedance of the (n-1)th layer plate. n Let be the thickness of the nth cavity layer.

[0009] Furthermore, the calculation method for the upper surface acoustic impedance of the (n-1)th layer is as follows: , Where: Z s,n-1’Let ρ be the acoustic impedance of the lower surface of the (n-1)th layer plate. n-1 Let c be the equivalent density of the (n-1)th layer. n-1 Let k be the velocity of sound in the (n-1)th layer of the plate. n-1 Let ρ be the wave number of the (n-1)th layer. n-1 Let t be the equivalent density of the (n-1)th layer. n-1 The thickness is the (n-1)th layer.

[0010] Furthermore, when the (n-1)th layer is a homogeneous material, ρ n-1 The density ρ of its material itself m c n-1 The speed of sound c corresponding to its material itself m k n-1 The wave number k corresponding to its material itself m ; When the (n-1)th layer is a plate with through holes, ρ n-1 Its equivalent density ρ dm c n-1 The velocity of sound in the plate, c dm k n-1 The wave number k of the plate dm ,and , , , Where: Φ n-1 Let ρ be the perforation rate of the (n-1)th layer. hn-1 K is the equivalent density of the holes in the (n-1)th layer. d Let be the bulk modulus of the (n-1)th layer.

[0011] Furthermore, , , Where: d n-1 Let K be the diameter of the perforation hole in the (n-1)th layer. m Let K be the bulk modulus of the material of the (n-1)th layer. hn-1 F is the equivalent bulk modulus of the holes in the (n-1)th layer plate. d As a correction factor, and , , Where: γ is the specific heat ratio of air, P0 is atmospheric pressure, κ is the thermal conductivity of air, and C p For the specific heat at constant pressure of air, Φ m ω represents the porosity of the material of the (n-1)th layer. d To convert the angular frequency, M d It is the shape factor, and , , Where: σ m Let a1 be the flow resistance of the material of the (n-1)th layer, and a1 be the hole spacing of the (n-1)th layer.

[0012] Furthermore, , , Where: Z s,1 Z is the acoustic impedance of the upper surface of the first layer plate. s,1’ denoted as , where h1 is the acoustic impedance of the lower surface of the first layer plate, h1 is the cavity depth of the first layer, t1 is the thickness of the first layer plate, d1 is the perforation diameter of the first layer plate, and Φ1 is the perforation rate of the first layer plate.

[0013] Furthermore, an equivalent fluid model of the JCA porous material consisting of a cavity and a sound-absorbing panel was established, and the equivalent density ρ of the Biot acoustic microparameter of the sound-absorbing panel material was calculated. m Equivalent speed of sound c m Equivalent wavenumber k m .

[0014] Furthermore, the equivalent density ρ m Equivalent speed of sound c m Equivalent wavenumber k m The specific calculation method is as follows: the equipment measures the sound absorption coefficient α of the model. m Then, calculate according to the following equation: , , , Where h is the cavity depth in the model, t is the sound-absorbing panel thickness in the model, and Z is the depth of the cavity in the model. s Z is the acoustic impedance of the upper surface of the sound-absorbing board. s’ The acoustic impedance of the lower surface of the sound-absorbing plate.

[0015] Furthermore, in calculating ρ m c m k m At that time, by changing the cavity depth h and the sound-absorbing plate thickness t in the model, different sound absorption coefficients corresponding to different models are obtained, and then a set of equations are formed for calculation.

[0016] Furthermore, a partition is provided inside the back panel to divide the internal space of the back panel into independent sound-absorbing chambers. One sound-absorbing unit is assembled in one sound-absorbing chamber, and each sound-absorbing unit is sealed and isolated from each other.

[0017] Furthermore, the sound-absorbing unit has four plates and four cavities. The third and second plates are both made of aluminum foam, and the third plate has through holes, while the second plate does not.

[0018] Furthermore, each layer is secured with snap-fit ​​fasteners before being assembled into the back cavity.

[0019] The beneficial effects of this invention are: The sound absorption coefficient of the entire structure can be calculated in advance during the design of the sound-absorbing metastructure sound barrier of the present invention. Therefore, the coefficient unit structure can be continuously adjusted to obtain the desired sound absorption effect. Thus, when designing the structure, it is not necessary to make an actual model. The desired structure can be obtained through theoretical calculation alone, which greatly reduces the design difficulty and cost. Attached Figure Description

[0020] Figure 1 This is an exploded view of the sound-absorbing metamorphic sound barrier structure in Example 1; Figure 2 This is an exploded view of the sound-absorbing unit in Example 1; Figure 3 This is a cross-sectional view of the sound-absorbing unit in Embodiment 1 (due to the choice of the cross-sectional position, some holes in the plates are not shown). Figure 4 This is a schematic diagram of the model used in calculating the acoustic microstructure parameters of aluminum foam material in Example 1. Figure 5 The actual sound absorption coefficient of the sound-absorbing unit in Example 1; Figure 6 This is a schematic diagram illustrating how the snap-fit ​​assembly of the various plates is achieved in Example 2. In the diagram: 1. Panel, 2. Sound-absorbing unit, 21. Back cavity, 22. Perforated plate, 23. Lower layer aluminum foam board, 24. Upper layer perforated aluminum foam board, 25. Facing perforated plate, 211. First cavity, 221. First layer board, 212. Second cavity, 222. Second layer board, 213. Third cavity, 223. Third layer board, 214. Fourth cavity, 224. Fourth layer board, 3. Partition, 4. Back panel, 5. Buckle. Detailed Implementation

[0021] The present invention will be further described below with reference to the accompanying drawings. The drawings are for illustrative purposes only, representing schematic diagrams rather than actual physical objects, and should not be construed as limiting the scope of this patent. To better illustrate the embodiments of the present invention, some components in the drawings may be omitted, enlarged, or reduced, and do not represent the actual dimensions of the product. It is understandable to those skilled in the art that some well-known structures and their descriptions may be omitted in the drawings. Example 1

[0022] A method for designing a sound-absorbing metamorphic sound barrier structure, such as Figure 1As shown, the sound-absorbing superstructure sound barrier structure in this embodiment includes a concave back plate 4 to prevent sound leakage and provide back protection, a partition 3 disposed inside the back plate 4 to divide the internal space of the back plate 4 into independent sound-absorbing chambers, sound-absorbing units 2 assembled in the sound-absorbing chambers, and a ventilated panel 1 assembled at the top for wind pressure resistance. There are multiple sound-absorbing units 2, and one sound-absorbing unit 2 is assembled in one sound-absorbing chamber.

[0023] like Figure 2 As shown, the sound-absorbing unit 2 includes a concave back cavity 21 and, from bottom to top, a perforated plate 22, a lower layer of aluminum foam board 23, an upper layer of perforated aluminum foam board 24, and a facing perforated plate 25 disposed within the back cavity 21. The perforated plate 22 and the facing perforated plate 25 are both made of sound-impermeable material, allowing sound to pass only through their perforations. The lower layer of aluminum foam board 23 and the upper layer of perforated aluminum foam board 24 are both sound-absorbing panels, allowing sound to pass through them. In this embodiment, although aluminum foam is a sound-absorbing material with good sound absorption, in practical applications, suitable materials can be selected and different combinations can be made according to sound absorption requirements. Furthermore, the number of sound-absorbing panels is not limited to two; different numbers can be used. Whether and how the perforations are set on the sound-absorbing panels can also be selected according to the actual situation. Figure 3 As shown, the bottom perforated plate 22 serves as the first layer 221, with a first cavity 211 below it. The lower layer of aluminum foam board 23 serves as the second layer 222, with a second cavity 212 below it. The upper layer of perforated aluminum foam board 24 serves as the third layer 223, with a third cavity 213 below it. The facing perforated plate 25 serves as the fourth layer 224, with a fourth cavity 214 below it. The overall sound absorption coefficient of this structure is then calculated to determine its sound absorption effect. The specific calculation process is as follows: The first step is to adopt, as follows Figure 4 The model shown calculates various acoustic parameters of the aluminum foam board. By changing the cavity depth h and the aluminum foam board thickness t in the model, the sound absorption coefficient α corresponding to each model is measured using instruments. m Then, substitute the values ​​into the following equations to form a system of equations, and calculate the equivalent density ρ of the aluminum foam board. m Equivalent speed of sound c m Equivalent wavenumber κ m : , , , Where: Z s Z is the acoustic impedance of the upper surface of the aluminum foam plate. s’ Let be the acoustic impedance of the lower surface of the aluminum foam board, j be the imaginary unit, ρ0 be the air density, c0 be the speed of sound in the air, and κ0 be the wave number of the air.

[0024] The second step is to calculate the acoustic impedance of the upper and lower surfaces of each plate in sequence.

[0025] For the first layer plate 221: , , Where: Z s,1 Z represents the acoustic impedance of the upper surface of the first layer plate 221. s,1’ denoted as , h1 is the acoustic impedance of the lower surface of the first layer plate 221, t1 is the depth of the first layer cavity 211, d1 is the perforation diameter of the first layer plate 221, Φ1 is the perforation rate of the first layer plate 221, η is the dynamic viscosity coefficient of air, J0 is the 0th order Bessel function, and J2 is the second order Bessel function.

[0026] For the second layer plate 222: , , Where: Z s,2 Z represents the acoustic impedance of the upper surface of the second layer plate 222. s,2’ h2 is the acoustic impedance of the lower surface of the second layer plate 222, h2 is the depth of the second layer cavity 212, and t2 is the thickness of the second layer plate 222.

[0027] For the third layer, 223, since this layer has a perforated structure, it is necessary to perform equivalent transformations on the acoustic parameters of this perforated aluminum foam board to obtain its equivalent density ρ. dm Speed ​​of sound c dm wave number k dm The specific calculation process is as follows: First, calculate the conversion angular frequency ω. d and shape factor M d : , , Where: Φ m σ represents the porosity of the aluminum foam board itself. m Φ is the flow resistance of the aluminum foam board itself, a1 is the hole spacing of the third layer board 223, Φ3 is the porosity of the third layer board 223, P0 is the atmospheric pressure, and d3 is the perforation diameter of the third layer board 223.

[0028] Calculate the equivalent bulk modulus K of the holes in the third layer plate 223. h3 and correction factor F d : , , Where: γ is the specific heat ratio of air, κ is the thermal conductivity of air, and C p The specific heat at constant pressure of air, K m This represents the bulk modulus of the aluminum foam board.

[0029] Calculate the equivalent density ρ of the holes in the third layer plate 223. h3 The bulk modulus K of the third layer plate 223 d : , , Then, calculate the equivalent density ρ of the third layer plate 223. dm Speed ​​of sound c dm wave number k dm : , , , Finally, the acoustic impedance Z on the upper surface of the third layer plate 223 is calculated. s,3 and the acoustic impedance Z of the lower surface of the third layer plate 223 s,3’ : , , Where: h3 is the depth of the third cavity 213, and t3 is the thickness of the third plate 223.

[0030] For the fourth layer plate 224: , , Where: Z s,4 Z represents the acoustic impedance of the upper surface of the fourth layer plate 224. s,4’ t4 is the acoustic impedance of the lower surface of the fourth layer plate 224, d4 is the diameter of the perforation of the fourth layer plate 224, Φ4 is the perforation rate of the fourth layer plate 224, and h4 is the thickness of the fourth layer cavity 214.

[0031] The third step is to calculate the sound absorption coefficient α of the entire sound-absorbing unit 2: .

[0032] The fourth step is to create a model and use instruments to measure the sound absorption coefficient of the entire sound absorption unit 2 for verification.

[0033] When t1, t2, t3, and t4 are 1.5mm, 5mm, 5mm, and 1.5mm respectively, h1, h2, h3, and h4 are 7mm, 70mm, 19mm, and 6mm respectively, d1, d3, and d4 are 4mm, 18mm, and 3mm respectively, and Φ1, Φ3, and Φ4 are 10%, 60%, and 30% respectively, the sound absorption performance curve is as follows: Figure 5 As shown in Figure A, it can be seen that within the frequency range of 200~4000Hz, the average sound absorption coefficient reaches 0.94, and the noise reduction coefficient (NRC) reaches 0.97, demonstrating excellent wide-range sound absorption performance. If a simplified design is implemented, removing the first layer 221 and changing h2, h3, and h4 to 74mm, 20mm, and 9.5mm respectively, while keeping other parameters unchanged, the result is as follows... Figure 5 As shown in Figure B, the sound absorption performance decreases in the 200~4000Hz range, but improves in the 100~200Hz range, making it suitable for lower frequency applications. Example 2

[0034] In this embodiment, as Figure 6 As shown, the panels in the sound-absorbing unit 2 are assembled together with snaps and then placed into the back cavity 21 as a whole to ensure the stability of the sound-absorbing unit 2.

[0035] The above embodiments are for illustrative purposes only and are not intended to limit the invention. Those skilled in the art can make various changes or modifications without departing from the spirit and scope of the invention. Therefore, all equivalent technical solutions should also fall within the protection scope of the invention, which should be defined by the claims.

Claims

1. A method for designing a sound-absorbing metamorphic sound barrier structure, characterized in that: The sound-absorbing superstructure sound barrier structure includes a sound-absorbing unit (2) assembled in the space enclosed by a back plate (4) and a front plate (1). The sound-absorbing unit (2) includes a back cavity (21) and a perforated plate (22) located at the bottom layer (i.e., the first layer plate (221), a perforated protective plate (25) located at the top layer (i.e., the nth layer plate), and a layer or more of sound-absorbing plates located between the first layer plate (221) and the nth layer plate (i.e., the second layer plate (222) to the (n-1)th layer plate). Each layer plate has a cavity below it. The sound absorption coefficient α of the entire sound-absorbing unit (2) is calculated according to the following formula: ， wherein: n is an integer greater than 2, Z s,n is the surface acoustic impedance of the facing perforated panel (25) at the uppermost layer of the sound absorption unit, i.e. the nth panel, and p0is the density of air, c0is the speed of sound in air, and: ， where: Z s,n’ is the surface acoustic impedance of the nth layer, j is the imaginary unit, k0is the wave number in air, t n is the thickness of the nth layer, d n is the perforation diameter of the nth layer, Φ n is the perforation ratio of the nth layer, ω is the angular frequency, η is the dynamic viscosity coefficient of air, J0is the 0th order Bessel function, and J2is the 2nd order Bessel function; and ， where: Z s,n-1 is the surface acoustic impedance of the top surface of the n-1 layer, h n is the thickness of the n-th cavity.

2. The sound-absorbing metamorphic sound barrier structure design method as described in claim 1, characterized in that: The calculation method for the upper surface acoustic impedance of the (n-1)th layer is as follows: ， where: Z s,n-1’ is the acoustic impedance of the nth-1 layer plate, p n-1 is the equivalent density of the nth-1 layer plate, c n-1 is the speed of sound in the nth-1 layer plate, k n-1 is the wave number of the nth-1 layer plate, p n-1 is the equivalent density of the nth-1 layer plate, t n-1 is the thickness of the nth-1 layer plate.

3. The sound-absorbing metamorphic sound barrier structure design method as described in claim 2, characterized in that: When the (n-1)th layer is a homogeneous material, ρ n-1 The density ρ of its material itself m c n-1 The speed of sound c corresponding to its material itself m k n-1 The wave number k corresponding to its material itself m ; When the (n-1)th layer is a plate with through holes, ρ n-1 Its equivalent density ρ dm c n-1 The velocity of sound in the plate, c dm k n-1 The wave number k of the plate dm ,and ， ， ， Where: Φ n-1 Let ρ be the perforation rate of the (n-1)th layer. hn-1 K is the equivalent density of the holes in the (n-1)th layer. d Let be the bulk modulus of the (n-1)th layer.

4. The sound-absorbing metamorphic sound barrier structure design method as described in claim 3, characterized in that: ， ， Where: d n-1 Let K be the diameter of the perforation hole in the (n-1)th layer. m Let K be the bulk modulus of the material of the (n-1)th layer. hn-1 F is the equivalent bulk modulus of the holes in the (n-1)th layer plate. d As a correction factor, and ， ， Where: γ is the specific heat ratio of air, P0 is atmospheric pressure, κ is the thermal conductivity of air, and C p For the specific heat at constant pressure of air, Φ m ω represents the porosity of the material of the (n-1)th layer. d To convert the angular frequency, M d It is the shape factor, and ， ， Where: σ m Let a1 be the flow resistance of the material of the (n-1)th layer, and a1 be the hole spacing of the (n-1)th layer.

5. The sound-absorbing metamorphic sound barrier structure design method as described in claim 4, characterized in that: ， ， Among them: Z s,1 Z is the acoustic impedance of the upper surface of the first layer plate. s,1’ denoted as , where h1 is the acoustic impedance of the lower surface of the first layer plate, h1 is the cavity depth of the first layer, t1 is the thickness of the first layer plate, d1 is the perforation diameter of the first layer plate, and Φ1 is the perforation rate of the first layer plate.

6. The sound-absorbing metamorphic sound barrier structure design method as described in claim 5, characterized in that: An equivalent fluid model of the JCA porous material consisting of a cavity and a sound-absorbing panel was established, and the equivalent density ρ of the Biot acoustic microparameter of the sound-absorbing panel material was calculated. m Equivalent speed of sound c m Equivalent wavenumber k m .

7. The sound-absorbing metamorphic sound barrier structure design method as described in claim 6, characterized in that: Equivalent density ρ m Equivalent speed of sound c m Equivalent wavenumber k m The specific calculation method is as follows: the equipment measures the sound absorption coefficient α of the model. m Then, calculate according to the following equation: ， ， ， Where h is the cavity depth in the model, t is the sound-absorbing panel thickness in the model, and Z is the depth of the cavity in the model. s Z is the acoustic impedance of the upper surface of the sound-absorbing board. s’ The acoustic impedance of the lower surface of the sound-absorbing plate.

8. The sound-absorbing metamorphic sound barrier structure design method as described in claim 7, characterized in that: In calculating ρ m c m k m At that time, by changing the cavity depth h and the sound-absorbing plate thickness t in the model, different sound absorption coefficients corresponding to different models are obtained, and then a set of equations are formed for calculation.

9. The sound-absorbing metamorphic sound barrier structure design method as described in claim 1, characterized in that: The back panel (4) has a partition (3) inside, which divides the internal space of the back panel (4) into independent sound-absorbing chambers. A sound-absorbing unit (2) is assembled in a sound-absorbing chamber, and each sound-absorbing unit (2) is sealed and isolated from each other.

10. The sound-absorbing metamorphic sound barrier structure design method as described in claim 1, characterized in that: The sound-absorbing unit (2) has four plates and four cavities. The third plate (223) and the second plate (222) are both made of aluminum foam. The third plate (223) has through holes, while the second plate (222) does not have through holes.

Images