Fan silencer system

Abstract

To provide a fan silencing system capable of muffling discrete, multiple-frequency, narrow-band sounds generated by a fan while maintaining the fan's airflow.SOLUTION: An embodiment of the present invention comprises a fan and an acoustic resonance structure, and the acoustic resonance structure is located within a near-field region of a sound generated by the fan.SELECTED DRAWING: Figure 1

Description

[Technical Field]

[0001] The present invention relates to a fan silencer system. [Background technology]

[0002] Fans are known to generate a very narrow, strong frequency sound depending on the number of blades and rotation speed, which has become a problem as noise.To reduce this noise, it has been proposed to place a silencer in the path of the airflow (wind) generated by the fan.

[0003] For example, Patent Document 1 describes a noise suppressor for equipment equipped with a heat source such as a light source lamp unit and an exhaust fan for exhausting heat from the heat source, in which an air guide member for exhaust air from the exhaust fan is arranged in a sealed manner from the air outlet side of the exhaust fan to the outside of the equipment, an elastic membrane that can vibrate due to sound waves generated by the exhaust fan is arranged on a peripheral wall portion of the air guide member facing the ventilation path in a position where it at least collides with the flow of exhaust air but does not block the flow of air in the exhaust direction, and an air chamber is formed behind the elastic membrane. The noise suppressor described in Patent Document 1 causes the air flow (wind) generated by the fan to hit the elastic membrane, causing the elastic membrane to vibrate, thereby converting sound energy into vibration energy and suppressing noise.

[0004] Furthermore, the use of a resonance type silencer has been proposed to reduce narrow band noise. For example, Patent Document 2 describes an electric blower that includes an impeller having multiple blades, an air guide having multiple stator vanes arranged around the impeller, an electric motor that drives a rotating shaft to which the impeller is fixed, a substantially cylindrical fan case that has an air intake port in the center for introducing airflow into the impeller and an exhaust port on the side and is fixed to the electric motor while containing the impeller and the air guide, a soundproof tube that has an exhaust port and is airtightly fixed to the fan case while containing the entire electric motor, a substantially cylindrical sound-absorbing means that has a recess with a predetermined width and depth on its circumference and is provided at a predetermined location on the surface of the electric motor, and a flexible thin film that is provided on the open end face of the recess in the sound-absorbing means. Patent Document 2 describes that sound of a specific frequency determined by the depth of the recess is absorbed by resonance. [Prior art documents] [Patent documents]

[0005] [Patent Document 1] JP 2001-142148 A [Patent Document 2] Patent Publication No. 2008-036065 Summary of the Invention [Problem to be solved by the invention]

[0006] In a configuration like Patent Document 1, where sound is reduced by directing the airflow (wind) generated by a fan to an elastic membrane to vibrate the membrane, the membrane needs to be positioned so that the wind directly hits it in order to vibrate it strongly, and therefore the membrane is positioned in a way that partially blocks the airflow path of the airflow generated by the fan, which causes a large pressure loss in the fan and reduces the airflow. Furthermore, in Patent Document 1, a large wind pressure is applied to the elastic membrane, so the characteristics of the elastic membrane change when the fan's air volume and wind pressure change. Therefore, it is not possible to utilize the resonance effect formed by the characteristics of the elastic membrane and the back air layer. Therefore, it is not possible to obtain a significant sound-dampening effect by targeting a specific frequency generated by the rotation of the fan, making it difficult to obtain a significant sound-dampening effect for the fan.

[0007] It is known that fan noise occurs discretely at multiple frequencies depending on the number of blades and the rotation speed. A resonance-type silencer such as that disclosed in Patent Document 2 silences sound of a single frequency that coincides with the resonance frequency of the resonance-type silencer, and is less effective at silencing sound in other frequency bands. Therefore, there has been a problem in that it is difficult to silence multiple discretely occurring sounds.

[0008] An object of the present invention is to provide a fan noise reduction system that solves the problems of the conventional technology described above and can reduce narrow-band noise of multiple discrete frequencies generated by a fan while ensuring the fan's airflow. [Means for solving the problem]

[0009] The present invention solves the problems by the following configuration.

[0010] [1] A fan and an acoustic resonance structure are included. A fan noise suppression system in which the acoustic resonance structure is located within the near-field region of the sound generated by the fan. [2] A fan noise suppression system according to [1], wherein the resonant frequency of the acoustic resonant structure corresponds to at least one frequency of the discrete frequency sound caused by the rotation of the fan blades. [3] A fan noise reduction system according to [1] or [2], wherein the area of ​​the acoustic resonance structure overlapping with the fan outlet when viewed from a direction perpendicular to the fan outlet is 50% or less of the area of ​​the fan outlet. [4] The fan noise silencing system according to any one of [1] to [3], wherein the acoustic resonance structure constitutes a part of the wall surface of the ventilation passage connected to the fan. [5] The fan noise silencing system according to any one of [1] to [4], wherein the surface provided with the vibrating body of the acoustic resonance structure is arranged parallel to an axis perpendicular to the air outlet of the fan. [6] A fan noise silencing system according to any one of [1] to [5], which has a sound-transmitting windbreak member on the side of the acoustic resonance structure that has the vibrating body. [7] The fan noise silencing system according to any one of [1] to [6], wherein the acoustic resonance structure is in contact with the fan. [8] The fan noise suppression system according to [7], wherein the acoustic resonance structure is in contact with the fan via a vibration-damping member. [9] A multi-layer acoustic resonant structure with different resonant frequencies; The fan noise reduction system according to any one of [1] to [8], wherein the acoustic resonance structure with a higher resonance frequency is positioned closer to the fan than the acoustic resonance structure with a lower resonance frequency.

[10] The fan noise reduction system according to any one of [1] to [9], wherein the acoustic resonance structure is disposed only downstream of the fan in the direction of airflow by the fan.

[11] A fan noise silencing system according to any one of [1] to [9], wherein the acoustic resonance structures are arranged upstream and downstream of the fan in the direction of airflow by the fan.

[12] A fan noise reduction system according to any one of [1] to

[11] , wherein the acoustic resonance structure is a membrane-type resonance structure having a membrane whose peripheral edge is fixed and supported so as to be capable of membrane vibration, and a back space formed on one side of the membrane.

[13] The fan noise reduction system according to

[12] , wherein the membrane-type resonant structure has a through hole that connects the back space with the outside.

[14] The fan silencing system according to any one of [1] to

[13] , wherein the fan is an axial fan. [Effects of the Invention]

[0011] According to the present invention, it is possible to provide a fan noise reduction system that can reduce narrow-band noise of a plurality of discrete frequencies generated by a fan while ensuring the air volume of the fan. [Brief explanation of the drawings]

[0012] [Figure 1] 1 is a perspective view schematically illustrating an example of a fan noise silencing system according to the present invention. [Figure 2] 2 is a view of the fan noise silencing system of FIG. 1 as seen from direction A. FIG. [Figure 3] FIG. 3 is a cross-sectional view of FIG. 2. [Figure 4] FIG. 10 is a cross-sectional view schematically showing another example of the fan noise silencing system of the present invention. [Figure 5] FIG. 10 is a cross-sectional view schematically showing another example of the fan noise silencing system of the present invention. [Figure 6] FIG. 10 is a cross-sectional view schematically showing another example of the fan noise silencing system of the present invention. [Figure 7] FIG. 10 is a cross-sectional view schematically showing another example of the fan noise silencing system of the present invention. [Figure 8] FIG. 10 is a cross-sectional view schematically showing another example of the fan noise silencing system of the present invention. [Figure 9] FIG. 10 is a cross-sectional view schematically showing another example of the fan noise silencing system of the present invention. [Figure 10] FIG. 10 is a cross-sectional view schematically showing another example of the fan noise silencing system of the present invention. [Figure 11] FIG. 10 is a cross-sectional view schematically showing another example of the fan noise silencing system of the present invention. [Figure 12] FIG. 10 is a cross-sectional view schematically showing another example of the fan noise silencing system of the present invention. [Figure 13] FIG. 1 is a diagram schematically illustrating a configuration of Comparative Example 1. [Figure 14] 1 is a graph showing the relationship between frequency and measured volume. [Figure 15] 1 is a graph showing the relationship between frequency and measured volume. [Figure 16] 1 is a graph showing the relationship between frequency and measured volume. [Figure 17] FIG. 10 is a diagram schematically illustrating the configuration of Comparative Example 2. [Figure 18] 10 is a graph showing the relationship between frequency and noise reduction volume. [Figure 19] 1 is a graph showing the relationship between frequency and measured volume. [Figure 20] 1 is a graph showing the relationship between frequency and measured volume. [Figure 21] 1 is a graph showing the relationship between frequency and measured volume. [Figure 22] 1 is a graph showing the relationship between frequency and measured volume. [Figure 23] 10 is a graph showing the relationship between frequency and noise reduction volume. [Figure 24] 1 is a graph showing the relationship between frequency and measured volume. [Figure 25] 1 is a graph showing the relationship between current and wind speed. [Figure 26] FIG. 10 is a diagram schematically illustrating the configuration of a fifth embodiment. [Figure 27] 1 is a graph showing the relationship between frequency and measured volume. [Figure 28] 1 is a graph showing the relationship between frequency and measured volume. [Figure 29] 1 is a graph showing the relationship between frequency and measured volume. [Figure 30] 1 is a graph showing the relationship between frequency and measured volume. [Figure 31] FIG. 12 is a diagram schematically illustrating the configuration of Comparative Example 7. [Figure 32] FIG. 13 is a diagram schematically illustrating the configuration of Example 9. [Figure 33] 1 is a graph showing the relationship between frequency and measured volume. DETAILED DESCRIPTION OF THE INVENTION

[0013] The present invention will be described in detail below. The following description of the components will be given based on a representative embodiment of the present invention, but the present invention is not limited to such an embodiment. In this specification, a numerical range expressed using "to" means a range that includes the numerical values ​​before and after "to" as the lower and upper limits. In this specification, the terms "orthogonal," "parallel," and "perpendicular" include the range of tolerance permitted in the technical field to which the present invention pertains. For example, "parallel" means being within a range of less than ±10° from the strict orthogonal direction, and the tolerance from the strict orthogonal direction is preferably 3° or less. The angle also means being within a range of less than ±10° from the strict angle. In this specification, the terms "same" and "match" include a margin of error generally accepted in the technical field.

[0014] [Fan silencer system] The fan noise suppression system of the present invention comprises: a fan and an acoustic resonance structure; An acoustically resonant structure is a fan noise suppression system that is located within the near-field region of the sound generated by the fan.

[0015] The near-field region of the sound generated by the fan is the region where the sound waves are in a near-field state. The near-field state of the sound waves is as follows. The propagation direction and intensity of sound waves generated from a sound source are eventually determined by differences in attenuation for each wave number and spatial constraints (duct walls, bends in the flow path, etc.). However, sound waves generated from a sound source are not governed by the influence of the above attenuation and constraints immediately after generation, and have amplitude over a wide range of wave numbers, including high wave number components that cannot propagate far. After propagating a certain distance, these sound waves become plane waves and their directionality is determined. The state immediately after generation of sound waves from this sound source is called the "near field" state. Therefore, the area near the sound source that satisfies the above conditions is called the near field region. According to wave theory, in this region, wave number components that cannot propagate farther become unable to propagate after propagating for about λ / 4. Therefore, since the fan, which is the sound source in the present invention, generates sound from the fan blades, the near-field region is the region at a distance of less than λ / 4 from the fan blades. If the fan is placed in the flow path, the near-field region is the region at a distance of less than λ / 4 from the fan along the flow path.

[0016] Sound in the near-field state (hereinafter also referred to as near-field sound) is the sound emitted from a sound source, including sounds that have higher wave numbers than the propagating sound waves and cannot propagate far (wave numbers where k > 2π × f / c, where c is the speed of sound and f is the frequency), and exists as if spatially clinging to the sound source. Specifically, in the wave equation that acoustic propagation follows, sound components with high wave numbers where k > 2π × f / c cannot propagate far from the sound source because the wave amplitude attenuates exponentially with distance, but in the near-field region, the effect of attenuation is small, so sounds with such high wave numbers blend together with the sound source and are localized only around the sound source as near-field sound.

[0017] In the fan noise reduction system of the present invention, it is believed that by arranging the acoustic resonance structure within the near-field region, the following two interactions occur with the near-field sound within the near-field region, thereby achieving a noise reduction effect. The first interaction mechanism is as follows: High-frequency sound waves in near-field sound are characterized by their small spatial wave size (the inverse of the wave number). This allows for spatially localized interactions with acoustic resonant structures placed near the sound source. Specifically, sound pressure is applied locally to only a small portion of the acoustic resonant structure. By creating such localized interactions in the acoustic resonant structure, which are difficult for sound waves with normal frequency waves that propagate far, it is easy to produce nonlinear effects in the acoustic resonant structure. The first interaction mechanism is thought to be that this nonlinear effect acts to dampen sounds of frequencies other than the target damping frequency (resonance frequency) of the acoustic resonant structure.

[0018] The second interaction mechanism is thought to be the effect of sound being reflected by the acoustic resonant structure and returning to the source, suppressing the generation of sound waves from the source. When a fan rotates, the blades cut through the air, generating tiny fluid vortices in the air around the blades. These vortices are deformed at the edges of the blades, generating sound; this is the mechanism by which fan noise (aerodynamic noise) is generated. By placing an acoustic resonance structure near the sound source, the sound generated from the sound source is reflected by the acoustic resonance structure, and this reflected sound propagates back to the sound source, interfering with the sound generated by the sound source. This interference reduces the sound pressure at the sound source. The effect of this is that the sound pressure at the sound source position decreases, which reduces the amount of sound emitted from the sound source, and this significantly reduces the radiated volume. Furthermore, it is highly likely that not only is the process of sound being emitted from the sound source suppressed, but the generation of the sound source itself, and in the case of this fan, the generation of microvortices themselves, is also suppressed. The acoustic resonance structure located in the near-field region interacts not only with sound waves emitted from the sound source that propagate far, but also with near-field sound that has a high wave number and remains near the sound source. By strongly interacting with this near-field sound with the acoustic resonance structure, the wave number mode of the sound emitted from the acoustic vortex is biased toward near-field sound, which is sound that does not propagate far. In addition, the reflection caused by this interaction reduces the sound pressure at the sound source position even in the near-field, extremely suppressing the generation of microvortices, which are the sound source. On the other hand, in an acoustic resonance structure placed in the far field, the sound pressure at the sound source position does not decrease at the near-field wave number, so the generation of microvortices that act as sound sources cannot be suppressed very much. Therefore, when an acoustic resonance structure is placed in the near-field region, which can cover the wave numbers of sound waves from low wave numbers to near-field sound wave numbers, the amount of microvortices that act as sound sources is extremely small. By reducing the amount of microvortices generated as sound sources, it is possible to reduce not only the frequency of the acoustic resonance structure but also aerodynamic noise of other frequencies. In particular, the peak noise of a fan is generated by the sound emitted from the microvortices of each blade being in phase with each other, resulting in a constructive interference effect and producing a strong sound. In this case, the energy is proportional to the square of the number of sound sources, so when the number of microvortices as sound sources decreases, the energy of the emitted sound decreases in proportion to the square of that number. Therefore, the sound is more susceptible to the sound reduction effect when the amount of microvortices generated decreases. Therefore, a selective noise reduction effect is produced for multiple peak noises. It is believed that the multiple discrete frequency noise suppression effect of the present invention is mainly due to the reduction in the number of sound sources through this second mechanism and the resulting peak noise suppression effect. In addition, noise other than fan peak noise, known as broadband noise (turbulent noise), is generated after the phases of the individual sound sources of the blades are out of sync and complex reinforcement and cancellation occurs. Therefore, it is thought that the noise volume will not decrease significantly even if the number of sound sources is reduced, and the result is that only the peak noise is selectively suppressed.

[0019] In the field of optics, for example, JR Lakowicz et al., "Radiative Decay Engineering: 2. Effects of Silver Island Films on Fluorescence Intensity, Lifetimes, and Resonance Energy Transfer," Analytical Biochemistry, 301, 261-277 (2002), shows the relationship between the distance between a metal particle and a fluorescent particle, the luminescence intensity, the lifetime of the light source, and the generation rate. It is thought that a similar phenomenon occurs with sound waves and sound sources. When an acoustic resonance structure is in the near-field region, the distance to the sound source is at most less than λ / 4, so the phase change of the sound wave due to propagation is small. However, the phase of the sound wave is inverted (a phase change of π) when reflected by the acoustic resonance structure. Therefore, the sound generated from the sound source and the sound reflected by the acoustic resonance structure and returned to the sound source have a phase shift that is almost inverted, so they interfere in antiphase. Therefore, the two sounds cancel each other out at the sound source position, generating a noise reduction effect at the sound source position.

[0020] As described above, the fan noise silencing system of the present invention can achieve a noise silencing effect over a wide frequency band regardless of the resonant frequency of the acoustic resonance structure by arranging the acoustic resonance structure in the near-field region, through a mechanism in which a nonlinear effect occurs due to localized interactions caused by spatially localized sound that is unique to near-field sound, and a mechanism in which the generation of fluid vortices, which are the sound source, is suppressed by reducing the sound pressure at the sound source position.As a result, a noise silencing effect can be achieved for sounds of multiple discrete frequencies (hereinafter also referred to as discrete frequency sounds) generated by fans.

[0021] Furthermore, the above two interaction mechanisms are the effects of the interaction between the sound source (sound waves) and the acoustic resonance structure, which is achieved by placing the acoustic resonance structure in the near-field region. Therefore, because wind flow is irrelevant, there is no need to place the acoustic resonance structure so that the wind directly hits the acoustic resonance structure. In other words, there is no need to place the acoustic resonance structure so that it partially blocks the airflow path generated by the fan. As a result, it is possible to muffle the noise generated by the fan while maintaining the fan's airflow volume.

[0022] As mentioned above, the near-field region is the region that is located at a distance of less than λ / 4 from the sound source. Therefore, the size of the near-field region varies depending on the wavelength (frequency) of the sound wave. In the present invention, when the resonant frequency of the acoustic resonance structure is fr (or the lowest order if there are multiple resonances), the wavelength is λ, and the area less than λ / 4 from the fan sound source is defined as the near-field area.

[0023] To achieve a higher noise reduction effect, it is preferable to position at least a portion of the acoustic resonance structure in an area at a distance of λ / 6 from the fan (sound source), and more preferably at a distance of λ / 8. The closer the distance between the sound source and the acoustic resonance structure, the smaller the phase change in the process of the sound being reflected by the acoustic resonance structure and returning to the sound source in the second mechanism described above, and therefore the higher the noise reduction effect due to interference between the reflected sound and the sound from the sound source.

[0024] In the present invention, an acoustic resonance structure resonates with sound waves at its resonance frequency to produce a sound-deadening effect. Various structures that produce a resonance phenomenon can be selected, but representative acoustic resonance structures include a membrane-type resonance structure, a Helmholtz resonance structure, and an air column resonance structure. Each acoustic resonance structure will be described in detail below.

[0025] The configuration of the fan noise silencing system of the present invention will be described with reference to the drawings. Fig. 1 is a schematic perspective view showing an example of a preferred embodiment of the fan noise silencing system of the present invention. Fig. 2 is a front view of Fig. 1 as seen from direction A. Fig. 3 is a cross-sectional view of Fig. 2. In Fig. 2, the acoustic resonance structure is shown in cross section. In Figs. 2 and 3, the fan rotor and other components are omitted, and only the outer shape and air outlet are shown.

[0026] The fan noise reduction system 10 shown in FIGS. 1 to 3 includes an axial fan 12a and a membrane-type resonant structure 30a.

[0027] The axial flow fan 12a is basically a known axial flow fan, and rotates a rotor having a plurality of blades to give kinetic energy to the gas, thereby blowing the gas in the axial direction. Specifically, the axial fan 12a has a casing 16, a motor (not shown) attached to the casing 16, and a rotor 18 having a shaft 20 attached to the motor and rotated, and blades 22 formed to protrude radially outward from the shaft 20. In the following description, the rotation axis of the shaft portion 20 (rotor 18) will be simply referred to as the "rotation axis," and the radial direction of the shaft portion 20 (rotor 18) will be simply referred to as the "radial direction."

[0028] The motor is a typical electric motor that rotates the rotor 18 .

[0029] The shaft portion 20 of the rotor 18 is substantially cylindrical, and one bottom surface side is attached to the rotating shaft of a motor, and is rotated by the motor. The blades 22 are formed on the circumferential surface of the shaft portion 20 so as to protrude radially outward from the circumferential surface. The rotor 18 has a plurality of blades 22, which are arranged in the circumferential direction of the circumferential surface of the shaft portion 20. In the example shown in FIG. 1, the rotor 18 is configured to have four blades 22, but this is not limited thereto and it is sufficient that the rotor has a plurality of blades 22. The casing 16 also has four frames in the drawing, but this is not limited thereto. The shape of the blades 22 may be any of the various shapes used in conventionally known axial flow fans.

[0030] Axial flow fan 12a generates airflow (wind) in the direction of the rotation axis when rotor 18 having blades 22 is rotated by a motor. There are no limitations on the direction of airflow, and the airflow may flow from the motor side in the direction of the rotation axis in the opposite direction to the motor, or from the opposite side to the motor toward the motor.

[0031] The casing 16 is fixed to the motor and radially surrounds the rotatable rotor 18 (blades 22). The thickness of the casing 16 in the direction of the rotation axis is greater than the thickness of the blades 22 and the shaft portion 20 so as to protect the rotor 18 from the outside.

[0032] Casing 16 has air outlet 16a that opens in the direction of the rotation axis, and rotor 18 is disposed inside air outlet 16a. When rotor 18, which has blades 22, rotates, air is drawn in from one opening side of air outlet 16a and sent out from the other opening side. In other words, the airflow (wind) generated by the rotation of rotor 18 is sent in the direction of the rotation axis.

[0033] The thickness of the casing 16 should be sufficient to protect the rotor 18 from the outside and to suppress the radial air flow generated by the rotation of the rotor 18, thereby increasing the amount of air flow in the direction of the rotation axis, and should be approximately 1.01 to 3.00 times the thickness of the blades 22 and / or shaft portion 20.

[0034] The axial flow fan 12a may further have various configurations that are included in known axial flow fans. For example, in the example shown in FIG. 1, the axial flow fan 12a has holes into which fastening members such as screws are inserted when fixing the axial flow fan 12a to various devices.

[0035] The membrane-type resonance structure 30a serves to muffle the discrete frequency sounds generated by the axial flow fan 12a. The membrane-type resonant structure 30a has a frame body 32 and a membrane 34, and is configured to form a back space 35 surrounded by the frame body 32 and the membrane 34, and resonates by membrane vibration of the membrane 34, which is supported so that it can vibrate on the frame body 32.

[0036] 1 to 3, the frame 32 has a rectangular parallelepiped shape with an opening having a bottom formed on one side thereof. That is, the frame 32 has a bottomed, rectangular tubular shape with one open side. The membrane 34 is a membrane-like member that covers the opening surface of the frame 32 where the opening is formed, and is supported so that it can vibrate with its peripheral edge fixed to the frame 32 . Furthermore, on the back side (frame body 32 side) of film 34, a back space 35 is formed that is surrounded by frame body 32 and film 34. In the example shown in Figs. 1 to 3, the back space is a closed space.

[0037] 1 to 3, the membrane-type resonance structure 30a is disposed downstream in the airflow direction of the axial fan 12a. The membrane-type resonance structure 30a is disposed in a position that does not block the airflow (air outlet 16a) from the axial fan 12a, specifically, around the area that serves as the airflow path for the air blown by the axial fan 12a. The membrane-type resonance structure 30a is disposed so that the membrane 34 is parallel to the rotational axis direction of the axial fan 12a (X direction in FIG. 3) and faces the rotational axis side.

[0038] Here, in the past, when an acoustic resonance structure such as a membrane-type resonance structure was used for sound attenuation, the resonance frequency of the acoustic resonance structure was adjusted to the frequency of the sound to be attenuated, and the sound of that frequency was attenuated by utilizing the resonance phenomenon. However, there was a problem that the effect of attenuating sounds in other frequency bands was low, and it was difficult to attenuate sounds of multiple discrete frequencies.

[0039] In contrast, in the fan noise silencing system of the present invention, by placing the membrane-type resonant structure 30a within the near-field region of the sound generated by the fan, the two interaction mechanisms described above are generated, making it possible to silence multiple discrete frequency sounds generated by the axial fan 12a. In this case, it is necessary that at least a part of the vibrating portion of the film 34 exists within the near-field region, and more preferably that the center of gravity of the vibrating portion of the film 34 exists within the near-field region.

[0040] Here, in the fan noise reduction system of the present invention, there is no particular limitation on the resonant frequency of the membrane-type resonant structure 30a (acoustic resonant structure). In addition, in order to effectively utilize the sound-dampening effect due to the inherent resonance of the acoustic resonance structure, it is desirable that the resonance frequency of the acoustic resonance structure be within the audible range (20-20,000 Hz), and more desirably in the range of 100-16,000 Hz. The resonant frequency of the membrane-type resonant structure 30a (acoustic resonant structure) preferably matches at least one of the discrete frequency sounds caused by the rotation of the fan blades, thereby enhancing the noise reduction effect for the discrete frequency sounds that match the resonant frequency of the acoustic resonant structure. For example, it is preferable that the resonant frequency of the acoustic resonance structure coincides with the discrete frequency sound with the greatest sound pressure, more specifically, the discrete frequency sound with the greatest A-weighted sound pressure level, among the discrete frequency sounds. This allows for effective muffling of the discrete frequency sound that contributes greatly to fan noise. It is also preferable that the resonant frequency of the acoustic resonance structure coincides with the lowest frequency sound among the multiple discrete frequency sounds. Because it is difficult to silence low frequencies with ordinary sound-deadening materials, the resonance effect can be used to selectively muffle low-frequency sounds, and then the structure can be combined with other sound-deadening materials. In the present invention, the resonance frequency of the acoustic resonance structure matching one of the discrete frequency sounds of the fan means that the resonance frequency of the acoustic resonance structure is within a range of ±10% of one of the discrete frequency sounds of the fan.

[0041] In the case of an axial fan, if the rotation speed is z (rps) and the number of blades is N, a strong sound (discrete frequency sound) is generated at a frequency of m×N×z (Hz) (m is an integer greater than or equal to 1). Furthermore, the resonant frequency of the membrane-type resonant structure is determined by the size (size of the vibrating surface, i.e., size of the opening of the frame 32), thickness, hardness, etc. of the membrane 34. Therefore, by adjusting the size, thickness, hardness, etc. of the membrane 34, the resonant frequency of the membrane-type resonant structure can be set appropriately.

[0042] As described above, the membrane-type resonant structure 30a has a back space 35 on the back side of the membrane 34. Because the back space 35 is closed, sound absorption occurs due to the interaction between the membrane vibration and the back space. Specifically, membrane vibration has frequency bands for the fundamental vibration mode and higher vibration modes, which are determined by the membrane conditions (thickness, hardness, size, fixing method, etc.), and which mode's frequency is strongly excited and contributes to sound absorption is determined by the thickness of the back space, etc. If the back space is thin, qualitatively it has the effect of making the back space harder, making it easier to excite higher vibration modes of membrane vibration.

[0043] 1 to 3, the back space 35 of the membrane-type resonant structure 30a is a closed space completely surrounded by the frame 32 and the membrane 34, but this is not limited to this and it is sufficient if the space is substantially partitioned so as to obstruct the flow of air, and in addition to being a completely closed space, the membrane 34 or the frame 32 may have a partial opening. Such a configuration with a partial opening is preferable because it can prevent changes in the sound absorption characteristics due to changes in the hardness of the membrane 34 caused by the gas in the back space expanding or contracting due to temperature changes, which applies tension to the membrane 34. By forming through-holes in the membrane 34, airborne sound propagation occurs. This changes the acoustic impedance of the membrane 34. The through-holes also reduce the mass of the membrane 34. These factors allow the resonant frequency of the membrane-type resonant structure 30a to be controlled. There is no particular limitation on the position where the through-hole is formed.

[0044] The thickness of the film 34 is preferably less than 100 μm, more preferably 70 μm or less, and even more preferably 50 μm or less. If the thickness of the film 34 is not uniform, the average value should be within the above range. On the other hand, if the film is too thin, it becomes difficult to handle. The film thickness is preferably 1 μm or more, and more preferably 5 μm or more. The Young's modulus of the film 34 is preferably 1000 Pa to 1000 GPa, more preferably 10000 Pa to 500 GPa, and most preferably 1 MPa to 300 GPa. The density of the membrane 34 is preferably 10 kg / m3 to 30,000 kg / m3, more preferably 100 kg / m3 to 20,000 kg / m3, and most preferably 500 kg / m3 to 10,000 kg / m3.

[0045] The thickness of the back space 35 (thickness in the direction perpendicular to the surface of the film 34) is preferably 10 mm or less, more preferably 5 mm or less, and even more preferably 3 mm or less. If the thickness of the rear space 35 is not uniform, the average value should be within the above range.

[0046] Furthermore, in the example shown in Figures 1 to 3, the shape of the membrane-type resonant structure 30a when viewed from a direction perpendicular to the surface of the membrane 34, i.e., the shape of the vibration region of the membrane 34, is rectangular, but this is not limited to this and may be circular, or may be a polygonal shape such as a triangular shape, an elliptical shape, etc.

[0047] As described above, in the fan noise reduction system of the present invention, the noise reduction effect is achieved by the interaction between the sound source (sound waves) and the acoustic resonance structure, which is achieved by arranging the acoustic resonance structure in the near-field region, and therefore the acoustic resonance structure does not need to be arranged so that the wind directly hits the acoustic resonance structure. From the perspective of ensuring the fan's airflow volume, it is preferable that the acoustic resonance structure be arranged so as not to block the airflow path of the airflow generated by the fan. Specifically, when viewed from a direction perpendicular to the fan's air outlet, the overlapping area between the acoustic resonance structure and the air outlet is preferably 50% or less of the area of ​​the air outlet, more preferably 10% or less, and even more preferably 0%, i.e., no overlap, as shown in Figure 2. Furthermore, if the acoustic resonance structure and the air outlet overlap, it is desirable to have a structure that allows the air to flow smoothly while suppressing the generation of wind noise, such as by installing a sloped structure.

[0048] Furthermore, it is preferable that the surface of the acoustic resonance structure that includes the vibrating body be arranged parallel to an axis that is perpendicular to the air outlet of the fan. In the example shown in Figure 2, the membrane 34 is the vibrating body of the membrane-type resonance structure 30a, and the surface of the membrane-type resonance structure 30a on which the membrane 34 is arranged is arranged parallel to the axis perpendicular to the air outlet 16a of the axial fan 12a. When the acoustic resonance structure is a Helmholtz resonance structure or an air column resonance structure, the air in the through-holes of the resonance structure is the vibrating body, and the surface on which the through-holes are formed is the surface on which the vibrating body is provided.

[0049] The wind from a fan is an unsteady fluid phenomenon, and when the unsteady wind strikes the membrane of a membrane-type resonance structure and vibrates it, the wind causes the membrane to vibrate. The vibrations generated in the membrane include a wide frequency spectrum, but at frequencies designed as the resonance of the membrane-type resonance structure, a resonant vibration phenomenon occurs on the membrane surface. In this resonant vibration, the vibration generated in the membrane tends to remain for a long time, and as the fan wind continues to flow, the resonant vibration tends to amplify. This can result in the membrane vibrating in this resonant vibration emitting sound, like a speaker. In particular, when a strong airflow is generated from the fan, if the resonance structure is positioned so that the wind from the fan strikes the membrane surface of the membrane-type resonance structure, the sound is amplified near the resonant frequency of the membrane-type resonance structure, and the noise reduction effect may not be achieved. Therefore, by configuring the surface having the vibrating body of the acoustic resonance structure to be arranged parallel to an axis perpendicular to the fan's air outlet, the air flow generated by the fan is prevented from hitting the surface having the vibrating body of the acoustic resonance structure and vibrating the membrane, thereby preventing the sound-absorbing effect from being reduced by wind.

[0050] Here, in the example shown in FIG. 1, the fan noise reduction system is configured to have one membrane-type resonance structure 30a (acoustic resonance structure), but this is not limited to this and the fan noise reduction system may be configured to have two or more acoustic resonance structures. For example, as shown in the example of FIG. 4, two membrane-type resonance structures 30a may be arranged downstream in the airflow direction of the axial flow fan 12a at positions that do not block the airflow (airflow opening 16a). In Figure 4, the two membrane-type resonance structures 30a are arranged so that the membranes 34 are parallel to the rotational axis direction of the axial fan 12a, the membranes 34 face the rotational axis side, and the membrane 34 sides of the two membrane-type resonance structures 30a face each other.

[0051] Furthermore, in the example shown in Fig. 4, the two membrane-type resonance structures 30a are arranged facing each other, but this is not limited to this, and the membrane-type resonance structures 30a may be arranged with their membrane surfaces flush and facing the same direction, as in the example shown in Fig. 5, the two membrane-type resonance structures 30a on the right side, the two upper membrane-type resonance structures 30a, and the two left membrane-type resonance structures 30a. Note that Fig. 5 is a view of the fan noise reduction system as seen from the direction of the rotation axis of the axial flow fan 12a, and the axial flow fan 12a is not shown.

[0052] Furthermore, when a fan is connected to the ventilation duct, the membrane-type resonance structure 30a (acoustic resonance structure) may form part of the wall surface (pipe 26) of the ventilation duct connected to the fan, as in the example shown in Figures 4 and 5. This allows the membrane-type resonance structure 30a to be positioned so as not to block the airflow (air outlet 16a).

[0053] In addition, in the example shown in Figure 1 etc., the membrane-type resonance structure 30a (acoustic resonance structure) is configured to be placed in a position directly in contact with the axial fan 12a (fan), but it may also be placed in a position separated from the fan as long as it is placed within the near-field region of the sound generated by the fan.

[0054] For example, in the example shown in Fig. 6, the membrane-type resonance structure 30b is disposed at a position separated from the axial fan 12a, and a duct 26 is disposed between the membrane-type resonance structure 30b and the axial fan 12a. That is, in the example shown in Fig. 6, the duct 26, which forms a passage for the airflow generated by the axial fan 12a, is connected to the downstream side of the axial fan 12a, and the membrane-type resonance structure 30b is disposed at the end of the outlet side of the duct 26.

[0055] From the viewpoint of placing the acoustic resonance structure within the near-field region of the sound generated by the fan, it is preferable that the acoustic resonance structure be placed in contact with the fan or along the outer periphery of the fan casing. If the acoustic resonance structure is a membrane-type resonance structure, it is preferable that the frame of the membrane-type resonance structure be in contact with the fan casing. The acoustic resonance structure and the fan may be fixed directly with screws or the like, or may be fixed via washers, or may be fixed via adhesive or pressure-sensitive adhesive.

[0056] Alternatively, the acoustic resonance structure is preferably disposed in contact with the fan via a vibration isolation member. 7, the side surface of the frame 32 of the membrane-type resonance structure 30a is in contact with the axial fan 12a via the vibration-damping member 36. By configuring the membrane-type resonance structure 30a to be in contact with the axial fan 12a via the vibration-damping member 36, it is possible to suppress transmission of vibrations from the axial fan 12a to the membrane-type resonance structure 30a, preventing the membrane of the membrane-type resonance structure 30a from vibrating due to the vibrations of the axial fan 12a and generating sound, and preventing the axial fan 12a and the membrane-type resonance structure 30a from resonating together.

[0057] The vibration-damping member 36 can be made of rubber, sponge, foam, or other materials commonly used as vibration-damping materials. Furthermore, by using the vibration-damping member also as a sound-absorbing material, such as a porous sound-absorbing material, it is possible to achieve both a broadband sound-absorbing effect at high frequencies and the suppression of vibration transmission to the resonating structure. Specifically, a foam-based sound-absorbing material such as Calmflex F2 manufactured by Inoac Corporation can be used.

[0058] Furthermore, when the fan noise reduction system has multiple acoustic resonance structures, it is preferable that the acoustic resonance structures have different resonance frequencies, since the fan noise reduction system has acoustic resonance structures with different resonance frequencies, thereby achieving a higher noise reduction effect against multiple discrete frequency sounds. 8, the fan noise reduction system includes a membrane-type resonant structure 30a and a membrane-type resonant structure 30b. The resonant frequency of the membrane-type resonant structure 30a is different from the resonant frequency of the membrane-type resonant structure 30b.

[0059] Here, when the fan noise reduction system has acoustic resonance structures with different resonance frequencies, it is preferable that the acoustic resonance structure with the higher resonance frequency is positioned closer to the fan than the acoustic resonance structure with the lower resonance frequency. 8, the resonant frequency of the membrane-type resonant structure 30a located closer to the axial fan 12a is higher than the resonant frequency of the membrane-type resonant structure 30b located farther from the axial fan 12a, thereby effectively reducing the number of discrete frequency sounds.

[0060] 1 and other examples, the acoustic resonance structure is arranged only downstream of the fan in the direction of airflow by the fan, but this is not limited to this, and the acoustic resonance structure may be arranged upstream of the fan, or may be arranged upstream and downstream of the fan as in the example shown in Fig. 9. In most devices, including server fans, it is desirable to be able to arrange the acoustic resonance structure in the space between the fan and the device case in order to reduce the noise heard by humans. In order to obtain a higher noise reduction effect, it is preferable that the acoustic resonance structure be disposed at least downstream of the fan, and it is more preferable that it be disposed upstream and downstream of the fan. When acoustic resonance structures are arranged on both the upstream and downstream sides of the fan, the resonance frequency of the upstream acoustic resonance structure and the resonance frequency of the downstream acoustic resonance structure may be the same or different.

[0061] Furthermore, the acoustic resonance structure may have a windbreak member that transmits sound on the side where the vibrating body is provided. Specifically, in the example shown in Figure 10, the fan noise reduction system has a membrane-type resonance structure 30a as an acoustic resonance structure, and a windbreak member 48 arranged on the surface of the membrane 34, which is the vibrating body of the membrane-type resonance structure 30a, covering the membrane 34. The windbreak member 48 is a member that allows sound to pass through but prevents wind from entering. By arranging the windbreak member 48 on the surface of the membrane 34, it is possible to prevent the airflow generated by the fan from exerting wind pressure on the membrane, which is the vibrating body of the membrane-type resonance structure, and thereby to prevent the membrane from vibrating, thereby preventing the sound deadening effect from being reduced by wind.

[0062] The windbreak member 48 can be made of a porous structure such as a foam such as sponge, especially an open-cell foam, or a fibrous material such as cloth or nonwoven fabric. It can also be made of a rubber material with an extremely low Young's modulus, such as silicone rubber, or a thin plastic film with a thickness of about 10 μm, such as cling film, which is loosely fixed rather than taut. These membranes differ significantly from the membrane 34 of the membrane-type resonance structure in terms of thickness, hardness, and fixing method, allowing sound to pass through without producing strong resonance in the audible range.

[0063] In addition, in the example shown in Figures 1 to 3, the fan noise silencing system is configured to have only the membrane-type resonant structure 30a, but this is not limited to this, and the fan noise silencing system may also be configured to further have a porous sound-absorbing material. For example, a porous sound-absorbing material may be provided in the space surrounded by the frame 32 and the membrane 34 of the membrane-type resonant structure 30a, i.e., in the back space 35. Alternatively, a porous sound-absorbing material may be provided on the surface of the membrane 34 of the membrane-type resonant structure 30a. By configuring the fan noise reduction system to include a porous sound-absorbing material, it is possible to reduce a wide frequency range of sounds other than the dominant sound that the resonator selectively reduces. The porous sound-absorbing material may also be used as a windbreak member.

[0064] The porous sound-absorbing material is not particularly limited, and any known porous sound-absorbing material can be used as appropriate. Examples of the porous sound-absorbing material include foamed materials and materials containing minute air particles, such as urethane foam, soft urethane foam, wood, sintered ceramic particles, and phenolic foam; glass wool, rock wool, microfiber (e.g., Thinsulate manufactured by 3M Co.); floor mats, carpets, melt-blown nonwoven fabrics, metal nonwoven fabrics, polyester nonwoven fabrics, metal wool, felt, insulation boards, and fiber and nonwoven fabric materials, such as glass nonwoven fabrics; wood wool cement boards; nanofiber materials, such as silica nanofibers; and gypsum boards.

[0065] There are no particular restrictions on the flow resistance of the porous sound-absorbing material, but it is preferably 1000 to 100,000 (Pa·s / m2), more preferably 3000 to 80,000 (Pa·s / m2), and even more preferably 5000 to 50,000 (Pa·s / m2). The flow resistance of porous sound absorbing materials can be evaluated by measuring the normal incidence sound absorption coefficient of a 1 cm thick porous sound absorbing material and fitting it with the Miki model (J. Acoust. Soc. Jpn., 11(1) pp.19-24 (1990)). Alternatively, it can be evaluated according to "ISO 9053". Also, a plurality of porous sound absorbing materials with different flow resistances may be stacked.

[0066] 1 to 3, the fan noise reduction system has a configuration including a membrane-type resonance structure 30a as the acoustic resonance structure, but is not limited to this. The fan noise reduction system may also have a Helmholtz resonance structure and / or an air column resonance structure as the acoustic resonance structure.

[0067] Fig. 11 shows a schematic cross-sectional view of an example of a fan noise suppression system having a Helmholtz resonance structure 40. The fan noise suppression system shown in Fig. 11 has the same configuration as the fan noise suppression system shown in Fig. 4, except that it has a Helmholtz resonance structure 40 instead of the membrane-type resonance structure 30a as the acoustic resonance structure.

[0068] 11 , the acoustic resonance structure is a Helmholtz resonance structure 40. The Helmholtz resonance structure 40 has a rectangular pillar-shaped frame 42 with an opening having a bottom on one side, and a plate-shaped lid 44 with a through hole 46 that covers the opening surface of the frame 32 and has its peripheral edge fixed to the frame 32. The Helmholtz resonance structure 40 is a structure in which air in an internal space 43 surrounded by the frame 42 and the lid 44 acts as a spring, and air in a through hole 46 formed in the lid 44 acts as a mass, causing resonance of the mass-spring and absorbing sound by thermoviscous friction in the vicinity of the wall of the through hole 46.

[0069] In the example shown in FIG. 11, the lid portion 44 having the through-hole 46 is arranged parallel to the direction of the rotation axis of the axial flow fan 12a, and the lid portion 44 faces the rotation axis side.

[0070] Conventionally, when a Helmholtz resonance structure is used for noise reduction, the resonance frequency of the Helmholtz resonance structure is adjusted to the frequency of the sound to be reduced, thereby reducing the frequency of the sound. However, this has resulted in a problem in that the reduction effect on sounds in frequency bands other than the resonance frequency is low, making it difficult to reduce the multiple discrete frequency sounds generated by fans.

[0071] In contrast, in the fan noise silencing system of the present invention, the Helmholtz resonator structure 40 is placed within the near-field region of the sound generated by the fan, thereby generating the two interaction mechanisms described above and making it possible to silence multiple discrete frequency sounds generated by the fan.

[0072] When the Helmholtz resonance structure 40 is used as the acoustic resonance structure, it is also preferable that the resonance frequency of the Helmholtz resonance coincides with one of the frequencies of the discrete frequency sounds generated by the axial flow fan 12a. The resonant frequency of the Helmholtz resonance is determined by the volume of the internal space surrounded by the frame body 42 and the lid portion 44 and the area, length, etc. of the through-hole 46. Therefore, by adjusting the volume of the internal space surrounded by the frame body 42 and the lid portion 44 of the Helmholtz resonance structure 40 and the area, length, etc. of the through-hole 46, the resonant frequency can be set appropriately.

[0073] 11, the through-hole 46 is formed in the lid 44, but this is not limiting, and the through-hole 46 may be formed in the frame 42. In this case, however, the entrance and exit of the through-hole must face the direction in which the discrete frequency sound generated by the axial fan 12a propagates, that is, the direction of the fan's flow path in FIG. In the example shown in FIG. 11, the Helmholtz resonance structure 40 is configured such that the frame body 42 and the lid portion 44 are separate bodies, but the frame body 42 and the lid portion 44 may be integrally formed.

[0074] In the Helmholtz resonance structure 40, the air inside the through-hole 46 is the vibrating body, and the surface of the lid 44 having the through-hole 46 is the surface that has the vibrating body. Therefore, it is preferable that the surface of the lid 44 having the through-hole 46 is arranged parallel to an axis perpendicular to the air outlet. Also, a windbreak member may be arranged on the surface of the lid 44.

[0075] Furthermore, the shape of the Helmholtz resonance structure 40 when viewed from a direction perpendicular to the surface of the lid portion 44 may be a square, or may be a polygonal shape such as a triangle, a circle, an ellipse, or the like.

[0076] 11, the fan noise reduction system is configured to have two Helmholtz resonance structures 40, but the present invention is not limited to this and may be configured to have one Helmholtz resonance structure, or may be configured to have three or more Helmholtz resonance structures. In the case of a configuration having multiple Helmholtz resonance structures, the frames of the Helmholtz resonance structures may be integrally formed, and further, the internal space may be shared. In addition, in the case of a configuration having a plurality of Helmholtz resonance structures, the Helmholtz resonance structures may have different resonance frequencies.

[0077] In the present invention, the resonator of the silencer may have an air column resonance structure. In the air column resonance structure, resonance occurs due to the generation of standing waves in a resonance tube with an opening.

[0078] Conventionally, when an air column resonance structure is used for noise reduction, the resonance frequency of the air column resonance structure is adjusted to the frequency of the sound to be reduced, thereby reducing the sound of that frequency. However, this has the drawback of being less effective at reducing the sound in frequency bands other than the resonance frequency, making it difficult to reduce the multiple discrete frequency sounds generated by fans.

[0079] In contrast, in the fan noise silencing system of the present invention, by placing an air column resonance structure within the near-field region of the sound generated by the fan, the two interaction mechanisms described above are generated, making it possible to silence multiple discrete frequency sounds generated by the fan.

[0080] When an air column resonance structure is used as the acoustic resonance structure, it is also preferable that the resonance frequency of the air column resonance matches one of the frequencies of the discrete frequency sounds generated by the fan. The resonant frequency of the air column resonance is determined by the length of the resonance tube, etc. Therefore, by adjusting the depth of the resonance tube, the size of the opening, etc., the frequency of the resonating sound can be set appropriately.

[0081] When an acoustic resonance structure has an internal space and a through-hole (opening) that connects the internal space to the outside, whether the resonance structure generates air column resonance or Helmholtz resonance depends on the size and position of the through-hole, the size of the internal space, etc. Therefore, by adjusting these appropriately, it is possible to select whether the resonance structure generates air column resonance or Helmholtz resonance. In the case of an air column resonance structure, a relatively wide opening is preferable because a narrow opening causes sound waves to be reflected at the opening, making it difficult for the sound waves to penetrate into the internal space. Specifically, if the opening is rectangular, the length of the short side is preferably 1 mm or more, more preferably 3 mm or more, and even more preferably 5 mm or more. If the opening is circular, the diameter is preferably within the above range. On the other hand, in the case of Helmholtz resonance, thermoviscous friction needs to occur in the through hole, so it is preferable that the through hole be relatively narrow. Specifically, if the through hole is rectangular, the length of the short side is preferably 0.5 mm to 20 mm, more preferably 1 mm to 15 mm, and even more preferably 2 mm to 10 mm. If the through hole is circular, the diameter is preferably within the above range.

[0082] The fan noise suppression system of the present invention may be configured to have different types of acoustic resonance structures, for example, a Helmholtz resonance structure and a membrane-type resonance structure. Here, from the viewpoint of miniaturization and thinning, it is preferable to use a membrane-type resonance structure as the acoustic resonance structure.

[0083] Materials for the frame and lid of the membrane resonance structure, Helmholtz resonance structure, and air column resonance structure (hereinafter collectively referred to as "frame materials") include metal materials, resin materials, reinforced plastic materials, and carbon fiber. Metal materials include aluminum, titanium, magnesium, tungsten, iron, steel, chromium, chromium molybdenum, nichrome molybdenum, copper, and alloys thereof. Resin materials include acrylic resin, polymethyl methacrylate, polycarbonate, polyamide-imide, polyarylate, polyetherimide, polyacetal, polyether ether ketone, polyphenylene sulfide, polysulfone, polyethylene terephthalate, polybutylene terephthalate, polyimide, ABS resin (acrylonitrile, butadiene, styrene copolymer synthetic resin), polypropylene, and triacetyl cellulose. Examples of reinforced plastic materials include carbon fiber reinforced plastics (CFRP) and glass fiber reinforced plastics (GFRP), as well as natural rubber, chloroprene rubber, butyl rubber, EPDM (ethylene propylene diene rubber), silicone rubber, and rubbers containing crosslinked structures of these. Various honeycomb core materials can also be used as the frame material. Honeycomb core materials are lightweight and highly rigid, so they are easily available as ready-made products. Honeycomb core materials made from a variety of materials can be used as the frame, including aluminum honeycomb cores, FRP honeycomb cores, paper honeycomb cores (manufactured by Shin-Nihon Feather Core Co., Ltd., Showa Aircraft Industry Co., Ltd., etc.), and thermoplastic resin (PP, PET, PE, PC, etc.) honeycomb cores (such as TECCELL, manufactured by Gifu Plastic Industry Co., Ltd.). Additionally, air-containing structures, such as foamed materials, hollow materials, and porous materials, can also be used as frame materials. To prevent airflow between cells when using multiple resonators, the frame can be formed using, for example, closed-cell foamed materials. Various materials can be selected, including closed-cell polyurethane, closed-cell polystyrene, closed-cell polypropylene, closed-cell polyethylene, and closed-cell rubber sponge. Closed-cell foams are more impermeable to sound, water, and gas than open-cell foams, and offer greater structural strength, making them suitable for use as frame materials. Furthermore, if the porous sound absorber described above has sufficient support, the frame can be formed solely from the porous sound absorber. Alternatively, the porous sound absorber and the frame materials can be combined, for example, by mixing or kneading. Using materials containing air inside the device can reduce device weight and provide thermal insulation.

[0084] Here, the frame material is preferably made of a material with higher heat resistance than a flame-retardant material, since it can be placed in a high-temperature location. Heat resistance can be defined, for example, by the time required to satisfy each item of Article 108-2 of the Building Standards Act Enforcement Order. A material that satisfies each item of Article 108-2 of the Building Standards Act Enforcement Order for 5 to 10 minutes is a flame-retardant material, a material that satisfies each item of Article 108-2 of the Building Standards Act Enforcement Order for 10 to 20 minutes is a quasi-noncombustible material, and a material that satisfies 20 minutes or more is a noncombustible material. However, heat resistance is often defined for each field. Therefore, depending on the field in which the fan silencer system will be used, the frame material should be made of a material with heat resistance equivalent to or higher than the flame-retardant defined in that field.

[0085] The thickness of the frame and the lid (frame thickness) is not particularly limited either, and can be set according to, for example, the size of the cross section of the opening of the frame.

[0086] Materials for the film 34 include various metals such as aluminum, titanium, nickel, permalloy, 42 alloy, Kovar, nichrome, copper, beryllium, phosphor bronze, brass, nickel silver, tin, zinc, iron, tantalum, niobium, molybdenum, zirconium, gold, silver, platinum, palladium, steel, tungsten, lead, and iridium; PET (polyethylene terephthalate), TAC (triacetyl cellulose), PVDC (polyvinylidene chloride), PE (polyethylene), PVC (polyvinyl chloride), PMP (polymethylpentene), COP (cycloolefin polymer), Zeonor, polycarbonate, etc. Resin materials such as acrylate, PEN (polyethylene naphthalate), PP (polypropylene), PS (polystyrene), PAR (polyarylate), aramid, PPS (polyphenylene sulfide), PES (polyethersulfone), nylon, PEs (polyesters), COC (cyclic olefin copolymers), diacetyl cellulose, nitrocellulose, cellulose derivatives, polyamide, polyamide-imide, POM (polyoxymethylene), PEI (polyetherimide), polyrotaxane (slide-ring material), and polyimide can be used. Glass materials such as thin-film glass and fiber-reinforced plastics such as CFRP (carbon fiber reinforced plastic) and GFRP (glass fiber reinforced plastic) can also be used. Rubbers such as natural rubber, chloroprene rubber, butyl rubber, EPDM, silicone rubber, and rubbers containing crosslinked structures of these can also be used. Combinations of these materials are also acceptable. When a metal material is used, the surface may be metal-plated to prevent rust.

[0087] In view of excellent durability against heat, ultraviolet rays, external vibrations, etc., it is preferable to use a metal material as the material for the film 34 in applications requiring durability.

[0088] Furthermore, the method for fixing the membrane or lid to the frame is not particularly limited, and methods using double-sided tape or adhesives, mechanical fixing methods such as screw fastening, pressure bonding, etc. can be used as appropriate. The fixing method, like the frame material and membrane, can be selected from the perspectives of heat resistance, durability, and water resistance. For example, adhesives such as Cemedine's "Super X" series, ThreeBond's "3700 Series (heat resistant)," and Taiyo Wire Mesh's "Duralco" series heat-resistant epoxy adhesives can be selected. Furthermore, double-sided tapes such as 3M's high-heat-resistant double-sided adhesive tape 9077 can be selected. In this way, various fixing methods can be selected to meet the required characteristics.

[0089] Here, in the example shown in FIG. 1 etc., the fan noise reduction system has an axial fan 12a as the fan, and is configured to suppress noise from an axial fan (propeller fan), but the present invention is not limited to this and can be applied to conventionally known fans such as sirocco fans, turbo fans, centrifugal fans, and line flow fans. A sirocco fan draws air in from the direction of the rotation axis of a rotor with blades and sends it out in a direction perpendicular to the rotation axis, and has an air outlet on the side. Therefore, for example, when the fan is a sirocco fan 12b as shown in Figure 12, a membrane-type resonance structure 30a (acoustic resonance structure) is arranged so as to be in contact with the air outlet 38. The configuration of the membrane-type resonance structure 30a is similar to the example shown in Figure 1 etc.

[0090] 12, the membrane-type resonance structure 30a is positioned so as not to block the air outlet of the sirocco fan 12b. The membrane-type resonance structure 30a is also positioned so that the membrane 34 is parallel to the direction perpendicular to the air outlet of the sirocco fan 12b and faces the air outlet.

[0091] Even in the case of a sirocco fan, sound is generated from the fan blades, so the area at a distance of less than λ / 4 from the fan blades is the near-field region. Therefore, by placing an acoustic resonance structure within the near-field region, the two interactions described above can be generated within the near-field region, resulting in a noise reduction effect. [Example]

[0092] The present invention will be described in more detail below with reference to examples. The materials, amounts used, ratios, treatment contents, treatment procedures, etc. shown in the following examples can be changed as appropriate without departing from the spirit of the present invention. Therefore, the scope of the present invention should not be construed as being limited by the examples shown below.

[0093] [Comparative Example 1] An axial fan (Model: 109P0612K701 manufactured by Sanyo Denki Co., Ltd.) was used. This axial fan has an outer diameter of 60 mm x 60 mm and a thickness of 15 mm. Because a casing is attached to the exhaust side of the fan, the distance from the front edge of the air outlet to the rotor blades is approximately 5 mm. To reduce the impact of solid vibrations from the fan, 5mm thick anti-vibration rubber is placed at the bottom of the fan. Additionally, to reduce the sound emitted from the sides of the fan as solid vibrations, the sides of the fan casing are surrounded by 5mm thick acrylic.

[0094] A square duct with an inner diameter of 60mm square, the same as the outer diameter of the fan, and a length in the duct direction of 30mm was created by cutting and combining rectangular plates with a short side length of 30mm from 5mm thick acrylic plates. The acrylic plates were processed using a laser cutter. This duct was placed on the surface of the fan's outlet side, matching the cross section of the fan's air path and the duct. The frame surrounding the fan casing and the outside of the duct were connected with tape to completely close the duct, creating a structure in which the duct is in close contact with the fan, as shown in Figure 13.

[0095] <Measurement> Using the fabricated structure, a fan was driven and the sound volume was measured. To measure the noise, a microphone (Ako 1 / 2 inch microphone 4152) was placed 200 mm axially away from the center of the fan, and 50 mm off the center axis in both the horizontal and vertical directions to avoid the influence of wind. The microphones were placed on both the exhaust and intake sides. The fan was driven using a DC stabilized power supply, with the fan driving conditions set at 12 V and 0.25 A.

[0096] The results of measurements taken with a microphone on the exhaust side are shown in Figure 14. The horizontal axis of the graph in Figure 14 is displayed logarithmically. Figure 14 shows that large peak noise (narrowband noise), which is a characteristic of fans with rotating blades, appears at multiple frequencies. In other words, it shows that discrete frequency noise is being generated. Among these, the large peaks are in integer multiple relationships. The volumes of 1.1 kHz and 2.2 kHz are particularly large.

[0097] Furthermore, the wind speed at the outlet end of the duct was measured using an anemometer and found to be 3.1 m / s. From then on, no change in wind speed was observed up to Example 3.

[0098] [Example 1] A fan noise reduction system was produced in the same manner as in Comparative Example 1, except that the inner wall of the duct was made of a membrane-type resonance structure produced as follows: The resonance frequency of the membrane-type resonance structure was set to 2.2 kHz.

[0099] <Design of membrane-type resonance structures> A membrane-type resonant structure was designed using acoustic-structural coupled calculations using the finite element method in COMSOL MULTIPHYSICS (COMSOL Inc.). The membrane material was PET, with a thickness of 75 μm, and the size and back-to-back distance were varied. A membrane-type resonant structure with a circular frame with an inner diameter of 24 mm, which is the vibrating part of the membrane, and a back-to-back distance of 6 mm was found to have a resonance at 2.2 kHz and high absorption. A back distance of 6 mm corresponds to a distance of 0.038 × λ for a wavelength of 2.2 kHz, λ, which shows that resonance can be achieved with an extremely thin structure. In the case of a normal air column resonance structure with a closed tube on one side, the required length is 0.25 × λ, so it can be seen that the thickness can be reduced to about 15% of the size of the air column resonance structure.

[0100] <Fabrication of membrane-type resonant structures> The structure designed above was created by processing an acrylic plate with a laser cutter. Specifically, a 3mm thick acrylic plate was processed to create two 30mm square perforated plate members with a 24mm diameter opening, and a 30mm square plate member. The two perforated plate members and the plate member were stacked in this order and attached with double-sided tape (Gemba no Chikara, manufactured by Askul) to create a frame.

[0101] A 75 μm thick PET film (Lumirror, manufactured by Toray Industries) was attached to the open surface of the frame with double-sided tape. By cutting the PET film to fit the outer shape of the frame, a membrane-type resonant structure was created with an outer shape of a 30 mm square, an inner frame shape of 24 mm, a PET film thickness of 75 μm, and a back-to-back distance of 6 mm. Six of these membrane-type resonant structures were fabricated, resulting in a duct (30 mm long) in which three of the four sides each had two membrane-type resonant structures (see Figure 5).

[0102] <Measurement> The fan of the fabricated fan noise reduction system was driven and the sound volume was measured on the exhaust side and the intake side in the same manner as in Comparative Example 1. The measurement results on the exhaust side are shown in Figure 15, and the measurement results on the intake side are shown in Figure 16. Figures 15 and 16 also show the results of Comparative Example 1.

[0103] From Figure 15, it can be seen that a large noise reduction effect of approximately 20 dB can be obtained at the resonant frequency of 2.2 kHz of the membrane-type resonant structure. Furthermore, it can be seen that a noise reduction effect can also be obtained for multiple discrete frequency sounds with different frequencies generated by fan rotation, as indicated by the arrows in Figure 15. In other words, it can be seen that a noise reduction effect can be obtained even for frequencies other than the resonant frequency of the membrane-type resonant structure. In this way, it can be seen that the fan noise reduction system of the present invention can reduce sounds with frequencies other than the resonant frequency of the acoustic resonant structure by arranging the acoustic resonant structure in the near-field region of the sound generated by the fan, thereby reducing the noise of multiple discrete frequency sounds with different frequencies generated by fan rotation. It is also clear that by matching the resonant frequency of the membrane-type resonant structure to one of the multiple discrete frequency sounds generated by fan rotation, the noise reduction effect at that frequency can be improved.

[0104] Furthermore, Figure 16 shows that the volume is reduced on the intake side as well, at the resonant frequency of the membrane-type resonance structure and at other frequencies. In other words, the sound silencing effect on the exhaust side does not simply reflect sound back to the intake side, but silencing both the exhaust and intake sides. This effect is thought to be due to the absorption of sound by the membrane-type resonance structure through membrane vibration, and the suppression of sound emission from the sound source due to interference between the sound reflected by the membrane-type resonance structure and the sound source.

[0105] In the fan noise reduction system of Example 1, the distance from the sound source (blade) of the fan to the center of the membrane vibrating part of the membrane-type resonance structure is 20 mm (5 mm distance from the front of the fan blade to the front of the air outlet) + 15 mm distance from the center position of the membrane of the membrane-type resonance structure to the front of the fan's air outlet). Since the wavelength / 4 of a frequency of 2.2 kHz is 39 mm, it can be seen that the membrane-type resonance structure is located within the near-field region.

[0106] Comparative Example 2 In Comparative Example 2, as shown in Fig. 17, the membrane-type resonance structure 30a was placed away from the axial fan 12a, and a duct 100 was placed between the membrane-type resonance structure 30a and the axial fan 12a. The membrane-type resonance structure 30a used was the same as the membrane-type resonance structure in Example 1. The duct 100 was the same as the duct in Comparative Example 1, except that its length was 60 mm. In this configuration, the distance between the sound source (blade) of the fan and the membrane-type resonant structure is 80 mm, so the membrane-type resonant structure 30a is located outside the near-field region.

[0107] <Measurement> The fan of the fan noise reduction system of Comparative Example 2 was driven and the volume was measured on the exhaust side and the intake side in the same manner as in Comparative Example 1. In Comparative Example 2, the volume was compared with the measurement results of the volume when the membrane-type resonance structure 30a was replaced with a duct, and the silencing volume was calculated from the difference. The results are shown in Figure 18. FIG. 19 shows the measurement results of the volume in Comparative Example 3 and in a case where the membrane-type resonant structure 30a of Comparative Example 3 is replaced with a duct (simple duct).

[0108] It can be seen from FIG. 18 that in Comparative Example 2, sound can be silenced at the resonant frequency of the membrane-type resonant structure 30a. However, from FIG. 19, which shows a wider frequency range, it can be seen that the structure of Comparative Example 3 does not provide a sound deadening effect at frequencies other than the resonant frequency of the membrane-type resonant structure 30a. In Comparative Example 2, the membrane-type resonant structure and the sound source are separated by λ / 2, so the sound silencing effect is achieved by the interference effect of normal sound waves (far-field interference). On the other hand, it is thought that the mechanism in the near-field region described above does not occur, so it is natural that it does not contribute to silencing sounds other than those at the resonant frequency of the membrane-type resonant structure.

[0109] In contrast, when a membrane-type resonant structure is placed in the near-field region as in Example 1, the interaction between the membrane-type resonant structure and the sound source must be treated as an integrated whole, and the interaction of high-wavenumber near-field sounds that do not propagate far must also be considered. In this case, the above-mentioned mechanism is thought to have contributed to the emission of sounds of frequencies other than the resonant frequency of the membrane-type resonant structure. Therefore, a sound silencing effect can be achieved for sounds over a wide frequency range in the near-field region.

[0110] From the above results, it can be seen that by arranging a membrane-type resonance structure in the near-field region, as in Example 1 of the present invention, it is possible to muffle multiple discrete frequency sounds generated by a fan. It can also be seen that by matching the resonant frequency of the membrane-type resonance structure with one of the discrete frequency sounds, a higher noise mitigation effect can be obtained at this frequency. It can also be seen that fan noise can be muffled without blocking the air path.

[0111] [Example 2] Using the same membrane-type resonance structure as in Example 2, a study was conducted in which the peak noise frequency was varied by changing the type of fan. A Sanyo Denki DC axial fan "9GA0612G9001" (frame size 60 mm, thickness 10 mm) was used. This fan was fixed in the same manner as in Example 1, and measurements were taken for a case in which the same membrane-type resonance structure as in Example 1 was attached to the exhaust side (Example 2), and a case in which a 30 mm duct with the same duct length as Example 1 was attached to the same position instead of a resonance structure (Comparative Example 3).

[0112] The measurement results are shown in Figure 20. In the case of this fan, the peak noise frequency appears at a frequency that is shifted from the resonant frequency of the membrane-type resonant structure. Around 2.2 kHz, which is the resonant frequency of the membrane-type resonant structure, sound attenuation of about 8 dB appears relatively widely. On the other hand, at the fan's peak noise frequencies (1.2 kHz, 2.4 kHz, 3.6 kHz), when the membrane-type resonant structure is in the near-field region, it can be seen that the sound is attenuated from the original peak volume. In this way, it can be seen that even for peak noise frequencies of the fan that are shifted from the resonant frequency of the membrane-type resonant structure, the peak noise can be silenced by the resonant structure in the near-field region. Regarding the amount of silencing of peak noise, it can be seen that in Example 1, where the resonant frequency is set to the fan peak noise frequency, the amount of silencing is greater and more preferable than in this example, where the resonant frequency is shifted from the fan peak noise frequency.

[0113] [Example 3] A membrane-type resonant structure was fabricated in the same manner as in Example 1, except that the resonant frequency of the membrane-type resonant structure was set to 1.1 kHz.

[0114] <Fabrication of membrane-type resonant structures> When designing using the finite element method with COMSOL MULTIPHYSICS, it was found that the resonant frequency of the membrane-type resonant structure of Example 1 could be increased to 1.1 kHz by increasing the back-to-back distance from 6 mm to 15 mm. This membrane-type resonant structure was fabricated using the same method as in Example 1, by processing an acrylic plate with a laser cutter.

[0115] The fabricated membrane-type resonant structure was placed 30 mm away from the surface of the fan's air outlet. A duct (pipe) was connected between the membrane-type resonant structure and the fan (see Figure 6). The distance from the center of the membrane-type resonant structure to the fan's sound source (blades) was 50 mm. On the other hand, since the wavelength / 4 of a frequency of 1.1 kHz is 78 mm, it can be seen that the membrane-type resonant structure was placed within the near-field region.

[0116] <Measurement> The fan of the fabricated fan noise reduction system was driven and the sound volume was measured on the exhaust side and the intake side in the same manner as in Example 1. The results are shown in Figure 21. Figure 21 also shows the measurement results of the volume when the membrane-type resonant structure of Example 3 was replaced with a duct (simple duct).

[0117] From Figure 21, it can be seen that a large noise reduction effect of about 10 dB can be obtained at the resonant frequency of 1.1 kHz of the membrane-type resonant structure. Furthermore, it can be seen that a noise reduction effect can be obtained even for multiple discrete frequency sounds generated by a fan.

[0118] [Example 4] A fan noise silencing system was fabricated in the same manner as in Example 1, except that the membrane-type resonance structure fabricated in Example 3 was placed downstream of the membrane-type resonance structure of the fan noise silencing system in Example 1 (see Figure 8). The results are shown in Figure 22. Figure 22 also shows the measurement results of the volume when the membrane-type resonant structure of Example 4 is replaced with a duct (simple duct).

[0119] It can be seen that a large noise reduction effect of about 15 dB can be obtained at the resonant frequencies of 1.1 kHz and 2.2 kHz for each membrane-type resonant structure. In other words, it can be seen that the noise reduction effect of each membrane-type resonant structure still works even when arranged in series. It can also be seen that a noise reduction effect can be obtained for multiple discrete frequency sounds generated by the fan, as indicated by the arrows in Figure 22. In other words, it can be seen that a noise reduction effect can be obtained for frequencies other than the resonant frequency of the membrane-type resonant structure.

[0120] The difference between the two sets of data in Figure 22 was taken and shown as the noise reduction level in Figure 23. It can be seen that near 1.1 kHz and 2.2 kHz, the fan noise peaks were reduced by more than 15 dB, and that a noise reduction effect was also obtained in other frequency bands.

[0121] To evaluate the noise level perceived by the ear for the fan noise suppression system of Example 4, octave band evaluation and overall noise level evaluation are shown. Figure 24 shows the results of evaluations for each 1 / 3 octave band, with the volume evaluated as an A-weighting evaluation (unit: dBA), which is a correction that takes into account the sensitivity of the human ear. By suppressing noise peaks at 1.1 kHz, 2.2 kHz, and other frequencies, it can be seen that overall noise is reduced, even in the 1 / 3 octave band evaluation, which evaluates by broadly averaging frequencies. Furthermore, A-weighting was applied to the entire audible frequency range, and the results were integrated to calculate the noise level. The noise level was 81.9 dBA for a simple duct, but was reduced to 74.9 dBA with the fan noise suppression system of Example 4. Since a difference of 3 dBA in noise level is considered sufficiently detectable by the average person, this 7 dBA noise suppression effect is a level that can be perceived as sufficiently quiet. In this way, we conducted research to suppress the discrete frequency sounds generated by fans, and showed that by placing an acoustic resonance structure within the near-field region, it is possible to silence not only the resonance frequency but also the entire discrete frequency sounds generated by the fan, thereby achieving a significant noise reduction effect.

[0122] [Example 5] In order to carry out measurements under stronger wind conditions than in Examples 1 to 4, the type of fan was changed. A 9GA0612P1J03 fan (thickness: 38 mm) manufactured by Sanyo Denki was used. Figure 25 shows the wind speed when the amount of current supplied to this fan was changed. By increasing the amount of current, a high wind speed and a high air volume can be obtained.

[0123] On the exhaust side of this fan, a membrane-type resonance structure with the same configuration as in Example 2 was placed. However, the membrane surface of the membrane-type resonance structure was lowered 5 mm toward the outer periphery than in Example 2 (see FIG. 26). This was done to allow for the placement of a windbreak member in Example 6, which will be described later.

[0124] <Measurement> The fan of the fabricated fan noise reduction system was driven and the sound volume was measured on the exhaust side and the intake side in the same manner as in Comparative Example 1. The measurement results on the exhaust side are shown in Figure 27. Also shown are the measurement results for Comparative Example 4, where the membrane-type resonance structure of Example 5 was replaced with a duct. The structural lengths in the flow path direction of Example 5 and Comparative Example 4 are both 30 mm, which is the same. Additionally, the wind speed at the outlet end of Example 4 and Comparative Example 4 was measured using an anemometer. The results were 14.5 m / s in both cases, confirming that there was no difference in wind speed between the case where the membrane-type resonance structure was attached and the case where the cylindrical structure was attached.

[0125] From Figure 27, it can be seen that a sound silencing effect can be obtained for peaks of frequencies other than the resonant frequency of the membrane-type resonant structure, as indicated by the arrows in Figure 27. However, for the peak around 1.1 kHz, which is the resonant frequency, it can be seen that there is an effect of amplifying the sound at frequencies around that frequency, and almost no peak silencing effect is obtained. In Example 5, the fan's airflow is large and unsteady because it is a rotating fan. This wind exerts wind pressure on the membrane surface, causing wind-induced vibrations on the membrane surface. The vibrations generated in the membrane include a wide frequency spectrum, but resonance occurs at frequencies designed as resonances in the membrane-type resonance structure, i.e., frequencies targeted for noise reduction, and their surrounding frequencies. At these resonance frequencies, the vibrations generated on the membrane surface tend to remain for a long time, and their amplitude tends to increase as the fan continues to operate. As a result, sound is emitted from there, just like a speaker. In this way, when a strong airflow is generated in the immediate vicinity of the fan, sound is amplified near the resonance frequency, and the desired noise reduction effect is thought to be largely unachieved.

[0126] [Example 6] A fan noise silencing system was fabricated in the same manner as in Example 5, except that a windbreak member was placed on the surface of the membrane of the membrane-type resonance structure (see FIG. 10). A urethane sponge (5 mm thick) was used as a windbreak. To minimize the effect on the membrane vibration, double-sided tape was not used on the membrane side of the sponge. Instead, Scotch tape was used to attach part of the air side of the sponge (the part that touches the frame of the membrane resonance structure below the sponge) to the side wall of the membrane resonance structure, preventing the sponge from slipping off the membrane resonance structure.

[0127] <Measurement> The fan of the fabricated fan noise reduction system was driven and the sound volume was measured on the exhaust side and the intake side in the same manner as in Comparative Example 1. The measurement results on the exhaust side are shown in Figure 28. The measurement results of Comparative Example 4 are also shown at the same time. Furthermore, the wind speed at the outlet end of Example 6 was measured using an anemometer, and the result was 14.5 m / s, confirming that the wind speed had not changed.

[0128] From Figure 28, it can be seen that the amplification of sound near the resonance frequency (1.1 kHz) that occurred in Example 5 can be significantly suppressed. Furthermore, as shown by the arrows in Figure 28, it can be seen that the effect of reducing peak sounds at frequencies other than the resonance frequency can also be obtained. Furthermore, from Figure 28, it can be seen that sound can be silenced over a wide band in the high frequency range above 5.4 kHz. This is due to the sound absorption effect of the sponge placed on the membrane surface. From the above results, it can be seen that by placing a windbreak material on the surface of the membrane, it is possible to significantly suppress the phenomenon of noise occurring near the resonance frequency when a membrane-type resonance structure is placed very close to the fan. Furthermore, it can be seen that by using a porous sound-absorbing material as the windbreak material, it is possible to achieve both the sound-absorbing effect of the porous sound-absorbing material and the sound-absorbing effect of the membrane-type resonance structure.

[0129] [Example 7] A fan noise reduction system was fabricated in the same manner as in Example 5, except that a Helmholtz resonance structure was used as the acoustic resonance structure. A Helmholtz resonance structure with a resonance frequency of 1.1 kHz was designed, with a through-hole length of 3 mm, a through-hole diameter of 4 mm, an internal space thickness of 12 mm, and an internal space diameter of 24 mm. The acrylic plate was processed with a laser cutter to create a Helmholtz resonance structure. A fan noise reduction system was fabricated in the same manner as in Example 5, with six Helmholtz resonance structure cells forming the duct wall surface.

[0130] Figure 29 shows the measurement results when the current supplied to the fan was 0.3 A. Also shown is the measurement result when a duct of the same length was installed instead of the Helmholtz resonance structure (Comparative Example 5). The wind speed was 5.5 m / s.

[0131] Figure 29 shows that even when a Helmholtz resonance structure is used as the acoustic resonance structure, it is possible to obtain a sound-damping effect against peak sounds at frequencies other than the resonance frequency. On the other hand, the amount of sound mitigation for the peak at 1.1 kHz, which is the resonance frequency, is slight, and sound is amplified around it. This is the effect of wind noise generated in the through-hole of the Helmholtz resonance structure, which resonates at the resonance frequency of the resonance structure, amplifying the sound and causing it to ring.

[0132] [Example 8] The volume was measured in the same manner as in Example 7, except that the current supplied to the fan was 1.3 A. The measurement results are shown in Figure 30. The measurement results are also shown for the case where a duct of the same length was installed instead of the Helmholtz resonance structure (Comparative Example 6). The wind speed was 15.1 m / s.

[0133] From Figure 30, it can be seen that the effect of silencing multiple peak sounds at frequencies other than the resonance frequency can be achieved even with a Helmholtz resonance structure under high airflow conditions. On the other hand, it can be seen that the wind noise amplified by resonance becomes louder as the wind speed increases, and that peak sounds near the resonance frequency are amplified. From the above, it can be seen that the effect of a resonant structure being able to silence multiple discrete frequency sounds is not limited to membrane-type resonators, but is a general effect. Furthermore, since the amplification effect of wind noise caused by Helmholtz resonance is greater than the phenomenon caused by membrane-type resonant structures, membrane-type resonant structures are considered preferable, especially when used in strong winds.

[0134] Comparative Example 7 To investigate application to fans other than axial fans, we investigated the application of this technology to a sirocco fan for blowers. We used a Sanyo Denki blower, model 9BMC12P2G001. This blower fan was placed on 10 mm thick vibration-isolating rubber, with air taken in from above and expelled horizontally. A 5 mm thick acrylic panel with an opening the same size as the air outlet (approximately 30 mm x 52 mm) was placed 30 mm away from the air outlet as a partition 102. The measurement microphone MP was positioned beyond this so that it was not directly exposed to the wind, and the experiment was conducted. The wind speed measured at the opening of the partition 102 was 7.7 m / s.

[0135] The measurements were carried out in a state where the air outlet and the opening of the partition 102 were connected by a duct 100 made of a 5 mm thick acrylic plate. A schematic diagram is shown in FIG.

[0136] [Example 9] A fan noise silencing system was fabricated in the same manner as in Comparative Example 7, except that four membrane-type resonant structures 30a of Example 4 were arranged in a duct shape between the air outlet and the opening of the partition 102 (see FIG. 32). The distance between the membrane-type resonant structure 30a and the blades of the sirocco fan is at least 24 mm, and the membrane-type resonant structure 30a is disposed within the near-field region.

[0137] <Measurement> In Example 9 and Comparative Example 7, the fan was driven and the volume was measured with the measuring microphone MP. The measurement results are shown in FIG.

[0138] 33, it can be seen that the configuration of Example 9 can reduce peak sounds near the resonance frequency, and also has a noise silencing effect on peak sounds that appear at other frequencies. This result shows that even with a sirocco fan, as with an axial flow fan, it is possible to achieve a noise silencing effect on multiple discrete frequency sounds by arranging an acoustic resonance structure in the near-field region. The above results clearly demonstrate the effectiveness of the present invention. [Explanation of symbols]

[0139] 10 Fan silencer system 12a axial fan 12b Sirocco Fan 16 Casing 16a Air outlet 18 rotor 20 Shaft 22 Feather 26 Conduit 30a, 30b Membrane type resonance structure 32, 42 frame 34 membrane 35 Back space 36 Vibration-proof member 38 Ventilation vent 40 Helmholtz resonance structure 43 Interior Space 44 Lid 46 Through hole 48 Windbreak material 100 duct 102 Screen MP Microphone

Claims

1. a fan and an acoustic resonance structure; the acoustic resonant structure is located within a near-field region of sound generated by the fan; a surface of the acoustic resonance structure that includes a vibrating body is parallel to an axis perpendicular to an air outlet of the fan and is parallel to a direction of a rotation axis of the fan; A fan noise reduction system in which the near-field region is within a distance range of less than λ / 4 from the blade portion of the fan, where λ is the wavelength of the resonant frequency of the acoustic resonant structure.

2. 2. The fan noise suppression system of claim 1, wherein the resonant frequency of the acoustic resonant structure corresponds to at least one frequency of a discrete frequency sound resulting from the rotation of the fan blades.

3. 3. The fan noise reduction system according to claim 1, wherein, when viewed from a direction perpendicular to the air outlet of the fan, an area where the acoustic resonance structure overlaps with the air outlet is 50% or less of an area of ​​the air outlet.

4. 4. The fan noise silencing system according to claim 1, wherein the acoustic resonance structure forms a part of a wall surface of an air passage connected to the fan.

5. 5. The fan noise suppression system according to claim 1, further comprising a sound-transmitting windbreak member on the side of said acoustic resonance structure on which said vibrating body is provided.

6. 6. The fan silencer system according to claim 1, wherein the acoustic resonance structure is in contact with the fan.

7. 7. The fan noise suppression system according to claim 6, wherein the acoustic resonance structure is in contact with the fan via a vibration-isolating member.

8. a plurality of said acoustic resonant structures having different resonant frequencies; 8. The fan noise reduction system according to claim 1, wherein the acoustic resonance structure with a higher resonance frequency is positioned closer to the fan than the acoustic resonance structure with a lower resonance frequency.

9. 9. The fan silencing system according to claim 1, wherein the acoustic resonance structure is disposed only downstream of the fan in the direction of airflow from the fan.

10. 9. The fan noise suppression system according to claim 1, wherein the acoustic resonance structures are arranged on the upstream side and downstream side of the fan in the direction of airflow by the fan.

11. The fan noise reduction system according to any one of claims 1 to 10, wherein the acoustic resonance structure is a membrane-type resonance structure having a membrane whose peripheral edge is fixed and supported so as to be capable of membrane vibration, and a back space formed on one side of the membrane.

12. The fan noise suppression system according to claim 11, wherein the membrane-type resonant structure has a through-hole that connects the rear space with the outside.

13. 13. A fan silencing system according to claim 1, wherein the fan is an axial fan.

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