阅读说明：本技术 风扇消音系统 (Fan silencing system ) 是由 白田真也 大津晓彦 山添昇吾 于 2020-03-24 设计创作，主要内容包括：提供一种风扇消音系统,其在能够确保风扇风量的同时,将风扇产生的离散的多个频率的窄频带的声音进行消音。风扇消音系统具有风扇及声共振结构,声共振结构配置在风扇产生的声音的近场区域内。(Provided is a fan silencing system which can ensure fan air quantity and simultaneously can silence the narrow-band sound of a plurality of discrete frequencies generated by a fan. The fan silencer system has a fan and an acoustic resonant structure disposed in a near field region of sound generated by the fan.)

1. A fan noise reduction system has a fan and an acoustic resonance structure,

the acoustic resonant structure is configured within a near field region of sound generated by the fan.

2. The fan silencer system of claim 1, wherein,

the resonant frequency of the acoustic resonant structure coincides with the frequency of at least one of the discrete frequency sounds caused by the rotation of the blades of the fan.

3. The fan silencer system according to claim 1 or 2, wherein,

an area of the acoustic resonance structure overlapping with the air blowing port is 50% or less with respect to an area of the air blowing port when viewed from a direction perpendicular to the air blowing port of the fan.

4. The fan muffler system according to any one of claims 1 to 3, wherein,

the acoustic resonance structure constitutes a part of a wall surface of an air duct connected to the fan.

5. The fan muffler system according to any one of claims 1 to 4, wherein,

the surface of the acoustic resonance structure having the vibrator is arranged in parallel to an axis perpendicular to the air blowing port of the fan.

6. The fan muffler system according to any one of claims 1 to 5, wherein,

the acoustic resonance structure has a wind-proof member that transmits sound on a surface side provided with the vibrator.

7. The fan muffler system according to any one of claims 1 to 6, wherein,

the acoustic resonant structure is in contact with the fan.

8. The fan silencer system of claim 7, wherein,

the acoustic resonance structure is in contact with the fan via a vibration-proof member.

9. The fan muffler system according to any one of claims 1 to 8,

the fan silencer system has a plurality of the acoustic resonance structures having different resonance frequencies,

the acoustic resonance structure having a high resonance frequency is disposed closer to the fan than the acoustic resonance structure having a low resonance frequency.

10. The fan silencer system according to any one of claims 1 to 9,

the acoustic resonance structure is disposed only on a downstream side of the fan in an air blowing direction by the fan.

11. The fan silencer system according to any one of claims 1 to 9,

the acoustic resonance structure is disposed on an upstream side and a downstream side of the fan in an air blowing direction by the fan.

12. The fan silencer system according to any one of claims 1 to 11,

the acoustic resonance structure is a membrane resonance structure having a membrane whose peripheral edge portion is fixed and which is supported so as to be capable of membrane oscillation, and a back surface space formed on one surface side of the membrane.

13. The fan silencer system of claim 12, wherein,

the film-type resonance structure has a through hole communicating the back surface space with the outside.

14. The fan silencer system according to any one of claims 1 to 13,

the fan is an axial fan.

Technical Field

The invention relates to a fan silencing system.

Background

It is known that a fan generates a strong sound in a very narrow frequency band at a frequency corresponding to the number of blades and the number of rotations thereof, and this sound becomes a problem as noise. In order to reduce such noise, it is proposed to dispose a silencer in a passage of an air flow (wind) generated by the fan.

For example, patent document 1 describes a silencer device in an apparatus including a heat source such as a light source lamp unit and an exhaust heat fan for heat exhaust of the heat source, the silencer device being configured to seal and dispose an air guide member for exhaust air of the exhaust fan from an air outflow side of the exhaust fan to an outside of the apparatus, wherein an elastic membrane body that can vibrate freely due to sound waves generated by the exhaust fan is disposed on a peripheral wall portion of the air guide member facing an air duct, at least at a position where the elastic membrane body collides with the exhaust air flow and does not close the air flow in an exhaust direction, and an air chamber is formed on a back surface side of the elastic membrane body. The silencer device described in patent document 1 vibrates the elastic membrane body by bringing an air flow (wind) generated by the fan into contact with the elastic membrane body, thereby converting sound energy into vibration energy to perform silencing.

It is proposed to use a resonance type muffler for reducing noise in a narrow frequency band.

For example, patent document 2 describes an electric blower including: an impeller having a plurality of blades; an air guide having a plurality of fixed blades arranged around an impeller; a motor for driving a rotating shaft to which an impeller is fixed; a substantially cylindrical fan casing having an air inlet at the center thereof for allowing an air flow to flow into the impeller, having an air outlet at a side surface thereof, and being fixed to the motor in a state of including the impeller and the air guide; a sound insulation tube having an air outlet and airtightly fixed to the fan housing in a state of containing the entire motor; a substantially cylindrical sound deadening member which is provided with a recess having a predetermined width and a predetermined depth on a circumference and is provided at a predetermined position on a surface of the motor; and a thin film portion provided on an opening end face of the recess of the noise reduction member and having flexibility. Patent document 2 describes that sound of a specific frequency determined according to the depth of the recess resonates and is attenuated.

Documents of the prior art

Patent document

Patent document 1: japanese patent laid-open No. 2001-142148

Patent document 2: japanese patent laid-open No. 2008 036065

Disclosure of Invention

Technical problem to be solved by the invention

As in patent document 1, in the case of a structure in which an air flow (wind) generated by a fan is brought into contact with an elastic film body to vibrate the elastic film body to perform noise reduction, it is necessary to arrange the wind so as to directly blow on the elastic film body in order to strongly vibrate the elastic film body, and therefore, the wind path is arranged so as to block a part of the air flow generated by the fan. Therefore, a large pressure loss is given to the fan, and the air volume is reduced.

In addition, in patent document 1, since a large wind pressure is applied to the elastic film body, the characteristics of the elastic film body change when the wind volume and the wind pressure of the fan change. Therefore, the resonance effect formed by the characteristics of the elastic film body and the back air layer cannot be utilized. Therefore, it is difficult to obtain a large sound deadening effect on the fan because an effect of greatly muffling a sound of a specific frequency generated by the rotation of the fan cannot be obtained.

It is known that noise of a fan is discretely generated at a plurality of frequencies corresponding to the number of blades and the rotational speed. The resonance type muffler as in patent document 2 muffles a sound of a single frequency that matches the resonance frequency of the resonance type muffler, and has a low muffling effect on sounds of other frequency bands. Therefore, there is a problem that it is difficult to mute sounds of a plurality of frequencies generated discretely.

The present invention has been made to solve the above-described problems of the prior art, and an object of the present invention is to provide a fan silencing system capable of silencing a narrow-band sound of a plurality of discrete frequencies generated by a fan while ensuring the fan air volume.

Means for solving the technical problem

The present invention solves the problem by the following configuration.

[1] A fan noise reduction system has a fan and an acoustic resonance structure,

the acoustic resonant structure is disposed in the near field region of the sound generated by the fan.

[2] The fan silencer system according to [1], wherein a resonance frequency of the acoustic resonance structure coincides with at least one of discrete frequency sounds caused by rotation of blades of the fan.

[3] The fan silencing system according to [1] or [2], wherein an area of the acoustic resonance structure overlapping the air blowing port is 50% or less with respect to an area of the air blowing port when viewed from a direction perpendicular to the air blowing port of the fan.

[4] The fan silencing system according to any one of [1] to [3], wherein the acoustic resonance structure constitutes a part of a wall surface of an air duct connected to the fan.

[5] The fan noise reduction system according to any one of [1] to [4], wherein a surface of the acoustic resonance structure provided with the vibration body is disposed in parallel to an axis perpendicular to an air outlet of the fan.

[6] The fan noise reduction system according to any one of [1] to [5], wherein a wind shielding member that transmits sound is provided on a surface side of the acoustic resonance structure where the vibration body is provided.

[7] The fan silencing system according to any one of [1] to [6], wherein the acoustic resonance structure is in contact with the fan.

[8] The fan muffling system according to item [7], wherein the acoustic resonance structure is in contact with the fan via a vibration-proof member.

[9] The fan muffler system according to any one of [1] to [8], which includes a plurality of acoustic resonance structures having different resonance frequencies,

the acoustic resonance structure having a high resonance frequency is disposed closer to the fan than the acoustic resonance structure having a lower resonance frequency.

[10] The fan muffling system according to any one of [1] to [9], wherein the acoustic resonance structure is disposed only on a downstream side of the fan in an air blowing direction by the fan.

[11] The fan muffler system according to any one of [1] to [9], wherein the acoustic resonance structure is disposed on an upstream side and a downstream side of the fan in an air blowing direction by the fan.

[12] The fan muffler system according to any one of [1] to [11], wherein the acoustic resonance structure is a film-type resonance structure having a film whose peripheral edge portion is fixed and which is supported so as to be able to vibrate the film, and a back space formed on one surface side of the film.

[13] The fan silencing system according to [12], wherein the film type resonance structure has a through hole communicating 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 flow fan.

Effects of the invention

According to the present invention, it is possible to provide a fan silencing system capable of silencing a narrow-band sound of a plurality of discrete frequencies generated by a fan while securing a fan air volume.

Drawings

Fig. 1 is a perspective view schematically showing an example of a fan silencing system according to the present invention.

Fig. 2 is a view of the fan silencing system of fig. 1 as viewed from direction a.

Fig. 3 is a cross-sectional view of fig. 2.

Fig. 4 is a cross-sectional view schematically showing another example of the fan silencer system of the present invention.

Fig. 5 is a cross-sectional view schematically showing another example of the fan silencer system according to the present invention.

Fig. 6 is a cross-sectional view schematically showing another example of the fan silencer system according to the present invention.

Fig. 7 is a cross-sectional view schematically showing another example of the fan silencer system of the present invention.

Fig. 8 is a cross-sectional view schematically showing another example of the fan silencer system according to the present invention.

Fig. 9 is a cross-sectional view schematically showing another example of the fan silencer system according to the present invention.

Fig. 10 is a cross-sectional view schematically showing another example of the fan silencer system according to the present invention.

Fig. 11 is a cross-sectional view schematically showing another example of the fan silencer system according to the present invention.

Fig. 12 is a cross-sectional view schematically showing another example of the fan silencer system according to the present invention.

Fig. 13 is a diagram schematically showing the structure of comparative example 1.

Fig. 14 is a graph showing a relationship between frequency and measured sound volume.

Fig. 15 is a graph showing a relationship between frequency and measured sound volume.

Fig. 16 is a graph showing a relationship between frequency and measured sound volume.

Fig. 17 is a diagram schematically showing the structure of comparative example 2.

Fig. 18 is a graph showing a relationship between frequency and sound deadening volume.

Fig. 19 is a graph showing a relationship between frequency and measured sound volume.

Fig. 20 is a graph showing a relationship between frequency and measured sound volume.

Fig. 21 is a graph showing a relationship between frequency and measured sound volume.

Fig. 22 is a graph showing a relationship between frequency and measured sound volume.

Fig. 23 is a graph showing a relationship between frequency and sound deadening volume.

Fig. 24 is a graph showing a relationship between frequency and measured sound volume.

Fig. 25 is a graph showing a relationship between current and wind speed.

Fig. 26 is a diagram schematically showing the structure of example 5.

Fig. 27 is a graph showing a relationship between frequency and measured sound volume.

Fig. 28 is a graph showing a relationship between frequency and measured sound volume.

Fig. 29 is a graph showing a relationship between frequency and measured sound volume.

Fig. 30 is a graph showing a relationship between frequency and measured sound volume.

Fig. 31 is a diagram schematically showing the structure of comparative example 7.

Fig. 32 is a diagram schematically showing the structure of example 9.

Fig. 33 is a graph showing a relationship between frequency and measured sound volume.

Detailed Description

The present invention will be described in detail below.

The following description of the constituent elements is made in accordance with a representative embodiment of the present invention, but the present invention is not limited to such an embodiment.

In the present specification, the numerical range expressed by the term "to" refers to a range including numerical values before and after the term "to" as a lower limit value and an upper limit value.

In the present specification, "orthogonal", "parallel" and "perpendicular" include error ranges that are allowable in the technical field to which the present invention pertains. For example, "parallel" means in a range of less than ± 10 ° with respect to strict orthogonality, and the like, and the error with respect to strict orthogonality is preferably 3 ° or less. Also, with respect to angles, it is also meant within a range of less than ± 10 ° with respect to the strict angle.

In the present specification, "the same" or "the same" includes an error range which is generally allowable in the technical field.

[ silencing System for Fan ]

The fan silencing system of the present invention is,

it has a fan and an acoustic resonance structure,

an acoustic resonant structure is disposed within a near field region of sound generated by the fan.

The near field region of the sound generated by the fan is a region in which the sound wave is in a near field state. In addition, the acoustic wave is as follows in the near-field state.

The acoustic waves generated from the acoustic source each determine the direction and intensity of propagation according to the attenuation difference of the wave number of each wave or the spatial restriction (the wall of the pipe, the curvature of the flow path, etc.). However, the sound wave generated from the sound source is not influenced by the above-described attenuation or restriction immediately after the sound wave is generated, and has an amplitude in a wide wavelength range including a high-frequency component which cannot propagate to a remote place. The acoustic wave becomes a plane wave after propagating over a certain distance, and the directivity is determined. The state immediately after the sound wave is generated from the sound source is referred to as a "near field" state. Accordingly, a region near the sound source satisfying the above condition is set as a near-field region.

In this region, it is known that, as a wave theory, a wave number component that cannot propagate far away cannot propagate during a propagation period of about λ/4.

Therefore, since the fan as a sound source in the present invention generates sound from the blade portion of the fan, the region less than λ/4 distance from the blade portion of the fan is a near-field region. In addition, when the fan is disposed in the flow path, a region along the flow path at a distance of less than λ/4 from the fan is a near field region.

The sound in the near-field state (hereinafter also referred to as near-field sound) includes a sound which has a higher wave number than the wave number of the propagating sound wave and cannot propagate to a far distance (wave number k > 2 pi × f/c when the sound velocity is c and the frequency is f) among the sounds emitted from the sound source, and exists in close spatial proximity to the sound source. Specifically, in the wave equation followed by sound propagation, a sound component with a high wave number of k > 2 π × f/c cannot propagate further than the sound source because the amplitude of the wave attenuates exponentially with respect to distance, whereas in the near-field region, such a sound with a high wave number is mixed with the sound source and exists only locally around the sound source as near-field sound because the influence of attenuation is small.

In the fan noise canceling system of the present invention, it is considered that the acoustic resonance structure is disposed in the near-field region, and the following two kinds of interactions are generated with respect to the near-field sound in the near-field region, thereby obtaining the noise canceling effect.

The mechanism of the first interaction is as follows.

The high-wavenumber acoustic wave of the near-field sound is characterized by a small size (reciprocal of wavenumber) of the spatial wave. Therefore, the acoustic resonance structure disposed close to the sound source can spatially interact locally. Specifically, the sound pressure is locally applied to only a very small portion of the acoustic resonance structure, and the like. By creating such local interactions in the acoustic resonant structure that are difficult for acoustic waves of typical wavenumbers propagating to far distances, nonlinear effects are easily created in the acoustic resonant structure. The mechanism of the first interaction is presumed that the silencing effect also acts on sounds of frequencies other than the target silencing frequency (resonance frequency) of the acoustic resonance structure by the nonlinear effect.

The second mechanism of interaction is presumed to be an effect of suppressing generation of sound waves from a sound source by sound returned to the sound source position by reflection by the acoustic resonance structure.

As the fan rotates, the blades cut the air, thereby creating a tiny vortex of fluid in the air surrounding the blades. The vortex is deformed at the edge of the blade or the like to generate sound, which is a mechanism of generating sound (aerodynamic sound) by the fan. By arranging the acoustic resonance structure in the vicinity of the sound source, the sound generated from the sound source is reflected by the acoustic resonance structure, and the reflected sound propagates to the sound source and interferes with the sound generated from the sound source. As a result of this interference, the sound pressure at the sound source position is reduced.

As an effect of this, first, the sound pressure at the sound source position decreases, and the amount of radiation of sound from the sound source decreases. This greatly reduces the radiated sound volume.

Further, there is a high possibility that not only the process of emitting sound from the sound source but also the generation of the sound source itself, and the generation of the minute vortex itself in the fan at this time can be suppressed. In the acoustic resonance structure disposed in the near-field region, not only the acoustic wave emitted from the sound source and propagating to a far distance but also the near-field sound having a high wave number and staying in the vicinity of the sound source interact. In the near-field sound, the wave number mode of the sound emitted from the acoustic eddy current is biased toward near-field sound, which is sound that does not propagate to a distant place, by strongly interacting with the acoustic resonance structure, and the sound pressure at the sound source position is reduced in the near-field by reflection due to the interaction, so that the generation amount of minute eddy current serving as a sound source is greatly suppressed.

On the other hand, in the acoustic resonance structure arranged in the far field, since the sound pressure at the sound source position does not decrease at the near-field wave number, the generation itself of the minute eddy current serving as the sound source cannot be suppressed almost. Therefore, when the acoustic resonance structure is disposed in a near-field region in which the wave number of the acoustic wave can be covered from a low wave number to the wave number of the near-field sound, the amount of generation of the minute eddy current serving as the sound source becomes extremely small.

Since the amount of generation of the minute eddy current which becomes the sound source is reduced, aerodynamic sound of other frequencies can be reduced as well as the frequency of the acoustic resonance structure. In particular, since the peak sound of the fan is in phase with the sound emitted from the minute vortices from the respective blades, interference effects are generated to reinforce each other and a strong sound is emitted. At this time, since the energy is proportional to the square of the number of sound sources, when the number of minute vortices as the sound source is reduced, the energy of the sound emitted as the square thereof is reduced. Therefore, the effect of reducing the sound when the amount of generation of the minute eddy current is reduced is likely to be greatly affected. Therefore, a selective sound deadening effect is exhibited for a plurality of peak sounds. It is considered that the plurality of discrete frequency sound suppression effects in the present invention are generated by mainly contributing to the reduction in the number of sound sources by the second mechanism and the peak sound suppression effect accompanying the reduction.

Further, since noise called broadband noise (turbulent noise) other than the fan peak sound is generated after the sound sources of the blades are generated in a complicated manner such that the phases are dispersed, mutually emphasized, and mutually canceled, it is considered that the noise amount is hardly reduced even if the number of sound sources is reduced, and as a result, only the peak sound is selectively suppressed.

This effect is shown in the optical field, for example, in JR Lakowicz et al, "radial Decay Engineering:2.Effects of Silver Island Films on Fluorescence Intensity, Lifetimes, and research Energy Transfer" Analytical Biochemistry,301,261 and 277(2002) for the distance of metal particles from fluorescent particles, the luminous Intensity or the lifetime and production rate of the light source. It is considered that the same effect is produced with respect to the sound wave or the sound source.

When the acoustic resonance structure is located in the near-field region, since the distance from the acoustic source is smaller than λ/4 at maximum, the phase change of the acoustic wave caused by propagation is small. On the other hand, the phase of the acoustic wave is inverted by reflection by the acoustic resonance structure (phase change of pi). Therefore, the sound generated from the sound source and the sound reflected by the acoustic resonance structure and returned to the sound source interfere with each other in the opposite phase because the phase shift is substantially in the phase inversion state. Therefore, the two kinds of sounds cancel each other at the sound source position, thereby producing a sound deadening effect at the sound source position.

As described above, the fan noise cancellation system according to the present invention can obtain a noise cancellation effect in a wide frequency band regardless of the resonance frequency of the acoustic resonance structure, based on the mechanism in which the nonlinear effect due to the local interaction is exhibited by the spatially local sound unique to the near-field sound by disposing the acoustic resonance structure in the near-field region, and the mechanism in which the generation of the fluid vortex as the sound source is suppressed by reducing the sound pressure at the sound source position. Therefore, a sound deadening effect is obtained for sounds of a plurality of discrete frequencies generated by the fan (hereinafter, also referred to as discrete-frequency sounds).

The two interaction mechanisms are effects of interaction between the acoustic source (acoustic wave) and the acoustic resonance structure, which are generated by disposing the acoustic resonance structure in the near-field region. Therefore, since it is not related to the flow of wind, it is not necessary to configure the acoustic resonance structure so that wind directly blows to the acoustic resonance structure. That is, there is no need to configure the acoustic resonance structure to block a portion of the duct of the air flow generated by the fan. Therefore, the sound generated by the fan can be silenced while ensuring the fan air volume.

Here, as described above, the region where the distance from the sound source is less than λ/4 is the near field region. Thus, the size of the near-field region differs depending on the wavelength (frequency) of the acoustic wave.

In the present invention, when the resonance frequency fr (the lowest order number when there are a plurality of resonances) of the acoustic resonance structure is set to λ, a region smaller than λ/4 from the fan sound source unit is set to a near-field region.

In addition, from the viewpoint of further improving the noise reduction effect, at least a part of the acoustic resonance structure is preferably disposed in a region at a distance of λ/6 from the fan (sound source), and more preferably in a region at a distance of λ/8. In the second mechanism described above, the closer the distance between the sound source and the acoustic resonance structure is, the smaller the phase change in the process of being reflected by the acoustic resonance structure and returning to the sound source becomes, and therefore the silencing effect by the interference of the reflected sound and the sound from the sound source is further improved.

In the present invention, the acoustic resonant structure resonates with the acoustic wave at its resonant frequency to produce a sound deadening effect. In the case of a structure in which a resonance phenomenon occurs, various structures can be selected, and for example, a membrane type resonance structure, a helmholtz resonance structure, and an air column resonance structure can be given as typical structures of the acoustic resonance structure. The respective acoustic resonance structures will be described in detail later.

The structure of the fan silencing system of the present invention will be explained with reference to the accompanying drawings.

Fig. 1 is a schematic perspective view showing an example of a preferred embodiment of a fan silencing system of the present invention. Fig. 2 is a front view of fig. 1 viewed from a direction a. Fig. 3 is a cross-sectional view of fig. 2. In addition, in fig. 2, the acoustic resonance structure is shown in cross section. In fig. 2 and 3, the rotor of the fan and the like are not shown, and only the outer shape and the air blowing port are shown.

The fan silencer system 10 shown in fig. 1 to 3 includes an axial fan 12a and a film resonant structure 30 a.

The axial flow fan 12a is basically a well-known axial flow fan that rotates a rotor having a plurality of blades to impart kinetic energy to the gas, thereby conveying the gas in an axial direction.

Specifically, the axial flow fan 12a includes a casing 16, a motor (not shown) attached to the casing 16, and a rotor 18 including a shaft 20 attached to the motor for rotation and blades 22 formed to protrude radially outward of the shaft 20.

In the following description, the rotation axis of the shaft 20 (rotor 18) is simply referred to as "rotation axis", and the radial direction of the shaft 20 (rotor 18) is simply referred to as "radial direction".

The motor is a normal electric motor that rotates the rotor 18.

The shaft portion 20 of the rotor 18 is substantially cylindrical, has one bottom surface side attached to a 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, and the plurality of blades 22 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 has a structure having 4 blades 22, but is not limited thereto as long as it has a plurality of blades 22. The number of frames of the case 16 is 4 in the drawing, but the present invention is not limited thereto.

The shape of the blade 22 can be various shapes used in conventionally known axial fans.

In the axial flow fan 12a, the rotor 18 having the blades 22 is rotated by a motor, thereby generating an air flow (wind) in the rotation axis direction. The flow direction of the air flow is not limited, and the air flow may flow from the motor side to the opposite side of the motor in the rotation axis direction, or may flow from the opposite side of the motor to the motor side.

The housing 16 holds the motor and surrounds the radial periphery of the rotatable rotor 18 (blades 22).

The thickness of the housing 16 in the rotation axis direction is larger than the thicknesses of the blades 22 and the shaft portion 20 so that the rotor 18 can be protected from the outside.

The casing 16 has an air outlet 16a opening in the rotation axis direction, and a rotor 18 is disposed in the air outlet 16 a. When the rotor 18 having the blades 22 rotates, air is taken in from one opening surface side of the air blowing port 16a, and air is taken in from the other opening surface side. That is, the air flow (wind) generated by the rotation of the rotor 18 is blown in the direction of the rotation axis.

The thickness of the casing 16 may be about 1.01 to 3.00 times the thickness of the blade 22 and/or the shaft 20, as long as the rotor 18 is protected from the outside and the airflow in the rotation axis direction is increased by suppressing the radial air flow in the air flow generated by the rotation of the rotor 18.

The axial flow fan 12a may have various structures as known axial flow fans.

For example, in the example shown in fig. 1, the axial flow fan 12a has a hole into which a fastening member such as a screw is inserted when the axial flow fan 12a is fixed to various apparatuses.

The film type resonance structure 30a muffles discrete frequency sound generated by the axial flow fan 12 a.

The film type resonance structure 30a has a structure in which a frame 32 and a film 34 are provided, and a back space 35 surrounded by the frame 32 and the film 34 is formed, and resonates by the film 34 supported by the frame 32 so as to be capable of vibrating.

In the example shown in fig. 1 to 3, the frame 32 has a rectangular parallelepiped shape and has an opening portion having a bottom surface formed on one surface. That is, the frame 32 has a bottomed square tube shape with one open surface.

The film 34 is a film-like member that covers the opening surface of the frame 32 where the opening is formed, and is supported so as to be vibratably fixed to the frame 32 at the peripheral edge portion.

A rear space 35 surrounded by the frame 32 and the film 34 is formed on the rear surface side (frame 32 side) of the film 34. In the examples illustrated in fig. 1 to 3, the back space is a closed space.

In the example shown in fig. 1 to 3, the film-type resonance structure 30a is disposed on the downstream side in the air blowing direction of the axial flow fan 12 a. The film-type resonance structure 30a is disposed at a position not blocking the air (the air blowing port 16a) blown by the axial flow fan 12a, specifically, around a region that becomes an air passage of the air blown by the axial flow fan 12 a. In the film-type resonance structure 30a, the film 34 is parallel to the rotational axis direction (X direction in fig. 3) of the axial fan 12a, and the film 34 is disposed toward the rotational axis side.

Here, conventionally, when an acoustic resonance structure such as a membrane type resonance structure is used for sound deadening, sound of a frequency to be deadened is deadened by matching the resonance frequency of the acoustic resonance structure with the frequency of the sound to be deadened and utilizing a resonance phenomenon. Therefore, the sound in other frequency bands is low in sound deadening effect, and there is a problem that it is difficult to deaden sounds of a plurality of discrete frequencies.

In contrast, in the fan noise canceling system of the present invention, the two interaction mechanisms are generated by disposing the film type resonance structure 30a in the near field region of the sound generated by the fan, so that the sound of a plurality of discrete frequencies generated by the axial flow fan 12a can be canceled.

At this time, at least a part of the vibratable portion of the film 34 needs to be present in the near field region, and more preferably, the position of the center of gravity of the vibratable portion of the film 34 needs to be present in the near field region.

Here, in the fan silencing system of the present invention, there is no particular limitation in the resonance frequency of the film type resonance structure 30a (acoustic resonance structure).

Further, in order to effectively apply the muffling effect by the original resonance of the acoustic resonance structure, the resonance frequency of the acoustic resonance structure is preferably in the audible range (20-20000Hz), more preferably in the range of 100-16000 Hz.

The resonance frequency of the film-type resonance structure 30a (acoustic resonance structure) is preferably coincident with at least one frequency of discrete-frequency sound caused by the rotation of the fan blades. This can further improve the sound deadening effect at a frequency matching the resonance frequency of the acoustic resonance structure in the discrete-frequency sound.

For example, the resonance frequency of the acoustic resonance structure preferably coincides with the magnitude of sound pressure in discrete-frequency sounds, more specifically, discrete-frequency sounds having the maximum a-characteristic sound pressure level. This makes it possible to effectively suppress a discrete frequency sound having a high degree of contribution to fan noise.

Further, the resonance frequency of the acoustic resonance structure preferably coincides with the lowest-frequency sound among the plurality of discrete-frequency sounds. Since the lower the frequency is, the more difficult the sound attenuation becomes in a typical sound attenuating material, it is possible to selectively attenuate low frequency sound by resonance effect and to combine it with another sound attenuating material.

In addition, in the present invention, the coincidence of the resonance frequency of the acoustic resonance structure with one of the discrete-frequency sounds of the fan is in a range in which the resonance frequency of the acoustic resonance structure is within ± 10% of the one of the discrete-frequency sounds of the fan.

In the case of an axial flow fan, when the number of rotations is z (rps) and the number of blades is N, sound (discrete frequency sound) is strongly generated at a frequency of m × N × z (hz) (m is an integer of 1 or more).

The resonance frequency of the film type resonance structure is determined by the size (the size of the vibration surface, that is, the size of the opening of the frame 32) of the film 34, the thickness, the hardness, and the like. Accordingly, the resonance frequency of the film type resonance structure can be appropriately set by adjusting the size, thickness, hardness, and the like of the film 34.

As described above, the film type resonance structure 30a has the rear space 35 on the rear surface side of the film 34. Since the back space 35 is closed, sound absorption is generated by the interaction of the film vibration with the back space.

Specifically, in the membrane vibration, there are a basic vibration mode and a frequency band of a higher-order vibration mode determined by the conditions (thickness, hardness, size, fixing method, and the like) of the membrane, and the frequency of which mode is strongly excited to contribute to sound absorption is determined by the thickness of the back space and the like. When the thickness of the back space is small, the effect of hardening the back space and the like are qualitatively produced, and thus a higher-order vibration mode of the film vibration is easily excited.

In the example shown in fig. 1 to 3, the back space 35 of the film-type resonance structure 30a is a closed space completely surrounded by the frame 32 and the film 34, but is not limited thereto, and may be a space substantially divided so that the flow of air is not obstructed, and may have a partial opening in the film 34 or the frame 32 in addition to the completely closed space. This manner of partially having the opening is preferable in the following respects: the gas in the back space expands or contracts due to the temperature change to apply tension to the film 34, and the hardness of the film 34 changes, thereby preventing the sound absorption characteristics from changing.

By forming the through-holes in the film 34, propagation by airborne sound occurs. This changes the acoustic impedance of the film 34. Further, the mass of the film 34 is reduced by the through-holes. This enables control of the resonance frequency of the film-type resonance structure 30 a.

The position where the through-hole is formed is not particularly limited.

The thickness of the film 34 is preferably less than 100 μm, more preferably 70 μm or less, and further preferably 50 μm or less. When the thickness of the film 34 is not uniform, the average value may be within the above range.

On the other hand, if the film thickness is thin, handling becomes difficult. The film thickness is preferably 1 μm or more, more preferably 5 μm or more.

The Young's modulus of the film 34 is preferably 1000Pa to 1000GPa, more preferably 10000Pa to 500GPa, and most preferably 1MPa to 300 GPa.

The density of the film 34 is preferably 10kg/m³～30000kg/m³More preferably 100kg/m³～20000kg/m³Most preferably 500kg/m³～10000kg/m³。

The thickness of the back space 35 (the thickness in the direction perpendicular to the surface of the film 34) is preferably 10mm or less, more preferably 5mm or less, and still more preferably 3mm or less.

When the thickness of the back space 35 is different, the average value may be within the above range.

In the examples shown in fig. 1 to 3, the shape of the film-type resonant structure 30a, that is, the shape of the vibration region of the film 34, as viewed in the direction perpendicular to the surface of the film 34 is a quadrilateral, but the shape is not limited to this, and may be a circle, a polygon such as a triangle, or an ellipse.

In the fan silencing system of the present invention, as described above, since the acoustic resonance structure is disposed in the near field region to produce an effect due to the interaction between the sound source (sound wave) and the acoustic resonance structure, the silencing effect can be obtained even if the acoustic resonance structure is disposed such that wind directly blows on the acoustic resonance structure. From the viewpoint of ensuring the fan air volume, the acoustic resonance structure is preferably arranged so as not to block the air passage of the air flow generated by the fan.

Specifically, when viewed in a direction perpendicular to the air blowing port of the fan, the area of the acoustic resonance structure overlapping the air blowing port is preferably 50% or less, more preferably 10% or less, and as shown in fig. 2, more preferably 0%, that is, not overlapping.

In addition, when the acoustic resonance structure and the air outlet overlap each other, a structure that suppresses the generation of wind noise while allowing wind to flow smoothly, such as a ramp-like structure, is desirably installed.

Preferably, the surface of the acoustic resonance structure on which the vibrator is provided is arranged parallel to an axis perpendicular to the air outlet of the fan.

In the example shown in fig. 2, the film 34 is a vibrating body of the film-type resonant structure 30a, and the surface of the film-type resonant structure 30a on which the film 34 is disposed in parallel with an axis perpendicular to the air blowing port 16a of the axial flow fan 12 a.

In addition, when the acoustic resonance structure is a helmholtz resonance structure or an air column resonance structure, the air in the through hole of the resonance structure is the vibrator, and the surface on which the through hole is formed is the surface on which the vibrator is provided.

The wind of the fan is an unstable fluid phenomenon, and if the unstable wind hits the membrane of the membrane type resonance structure and shakes the membrane, vibration due to the wind is generated on the membrane. The vibration generated on the membrane includes a broad frequency spectrum, but a resonance vibration phenomenon occurs on the membrane face at a frequency designed as a resonance of the membrane type resonance structure. In this resonance vibration, the vibration generated in the film tends to remain for a long time, and the resonance vibration tends to be amplified in the wind-continuous flow of the fan. This sometimes causes sound to be transmitted from the resonant diaphragm as a speaker. In particular, when the resonance structure is disposed so that the wind from the fan hits the membrane surface of the membrane type resonance structure under the condition that a strong wind amount is generated from the fan, sound is amplified near the resonance frequency of the membrane type resonance structure, and a noise cancellation effect may not be obtained.

Accordingly, by the configuration in which the surface provided with the vibrator having the acoustic resonance structure is arranged in parallel to the axis perpendicular to the air outlet of the fan, it is possible to suppress the air flow generated by the fan from hitting the surface provided with the vibrator having the acoustic resonance structure to cause the film to shake, and to suppress the reduction of the noise reduction effect due to the wind.

Here, in the example shown in fig. 1, the fan muffler system has one membrane-type resonance structure 30a (acoustic resonance structure), but the present invention is not limited to this, and may have two or more acoustic resonance structures.

For example, as illustrated in fig. 4, two membrane type resonance structures 30a may be disposed at positions on the downstream side in the air blowing direction of the axial flow fan 12a where the air blowing (air blowing port 16a) is not blocked.

In fig. 4, the two film type resonance structures 30a are arranged such that the film 34 is parallel to the rotational axis direction of the axial flow fan 12a, and the film 34 is directed toward the rotational axis side, and the surfaces of the two film type resonance structures 30a on the film 34 side face each other.

Further, in the example shown in fig. 4, the two film type resonance structures 30a are disposed so as to face each other, but the present invention is not limited to this, and the film type resonance structures 30a may be disposed in the same orientation with the film surfaces on the same plane, as the two film type resonance structures 30a on the right side, the two film type resonance structures 30a on the upper side, and the two film type resonance structures 30a on the left side in fig. 5 shown in fig. 5. Fig. 5 is a view of the fan noise reduction system as viewed from the direction of the rotation axis of the axial flow fan 12a, and the axial flow fan 12a is not shown.

In the case where the fan is connected to the air duct, as illustrated in fig. 4 and 5, the membrane-type resonance structure 30a (acoustic resonance structure) may constitute a part of the wall surface (duct 26) of the air duct connected to the fan. Thus, the membrane resonance structure 30a can be configured to be disposed at a position not blocking the air blow (the air blowing port 16 a).

In the example shown in fig. 1 and the like, the membrane type resonance structure 30a (acoustic resonance structure) is a structure of a pimple at a position directly contacting the axial flow fan 12a (fan), but may be disposed at a position apart from the fan if disposed in a near field region of the sound generated by the fan.

For example, in the example shown in fig. 6, the film-type resonance structure 30b is disposed at a position separated from the axial flow fan 12a, and the duct 26 is disposed directly between the film-type resonance structure 30b and the axial flow fan 12 a. That is, in the example shown in fig. 6, a duct 26 forming a passage for wind generated by the axial flow fan 12a is connected to the downstream side of the axial flow fan 12a, and a film-type resonance structure 30b is disposed at an end portion on the outlet side of the duct 26.

From the viewpoint of disposing the acoustic resonance structure in the near-field region of the sound generated by the fan, the acoustic resonance structure is preferably disposed in contact with the fan or along the outer periphery of the fan casing. In the case where the acoustic resonance structure is a membrane type resonance structure, the frame of the membrane type resonance structure is preferably in contact with the casing of the fan. The acoustic resonance structure and the fan may be directly fixed by screws or the like, may be fixed via a gasket, or may be fixed via an adhesive or an adhesive.

Alternatively, the acoustic resonance structure is preferably arranged to be in contact with the fan via a vibration isolation member.

In the example shown in fig. 7, the side surface of the frame 32 of the membrane-type resonance structure 30a is in contact with the axial flow fan 12a via the vibration isolation member 36. The membrane-type resonance structure 30a is configured to contact the axial flow fan 12a via the vibration isolation member 36, thereby suppressing the transmission of the vibration of the axial flow fan 12a to the membrane-type resonance structure 30a, and preventing the membrane of the membrane-type resonance structure 30a from vibrating by the vibration of the axial flow fan 12a to generate sound and the resonance in which the axial flow fan 12a and the membrane-type resonance structure 30a are integrated.

As the vibration-proof member 36, a member generally used as a vibration-proof member made of rubber, sponge, foam, or the like can be used. The vibration isolation member also serves as a sound absorbing material, for example, a porous sound absorbing material, and thus can simultaneously provide a broad-band sound absorbing effect at high frequencies and suppress transmission of vibration to the resonant structure. Specifically, a foam-based sound absorber such as CalmFlexF2 manufactured by Inoac Corporation can be used.

When the fan noise reduction system has a plurality of acoustic resonance structures, it is preferable to have acoustic resonance structures having different resonance frequencies. Since the fan noise reduction system has the acoustic resonance structure with different resonance frequencies, a higher noise reduction effect can be obtained for a plurality of discrete frequency sounds.

For example, in the example shown in fig. 8, the fan silencing system has a film type resonance structure 30a and a film type resonance structure 30 b. The resonance frequency of the film type resonance structure 30a is different from the resonance frequency of the film type resonance structure 30 b.

Here, when the fan muffler system has an acoustic resonance structure having different resonance frequencies, the acoustic resonance structure having a higher resonance frequency is preferably disposed closer to the fan than the acoustic resonance structure having a lower resonance frequency.

In the example shown in fig. 8, the resonance frequency of the film-type resonance structure 30a disposed on the side close to the axial flow fan 12a is higher than the resonance frequency of the film-type resonance structure 30b disposed on the side away from the axial flow fan 12 a. This can significantly reduce the noise of a plurality of discrete frequency sounds.

In the example shown in fig. 1 and the like, the acoustic resonance structure is arranged only on the downstream side of the fan in the air blowing direction by the fan, but the present invention is not limited to this, and the acoustic resonance structure may be arranged on the upstream side of the fan, and as illustrated in fig. 9, the acoustic resonance structure may be arranged on the upstream side and the downstream side of the fan. In most devices including server fans, it is desirable to be able to arrange an acoustic resonance structure in the space between the fan and the device housing in order to reduce noise heard by humans.

From the viewpoint of obtaining a higher noise reduction effect, the acoustic resonance structure is preferably disposed at least on the downstream side of the fan, and more preferably on the upstream side and the downstream side of the fan.

In the case where the acoustic resonance structure is disposed on the upstream side and the downstream side 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.

Further, a wind-proof member that transmits sound may be provided on the surface of the acoustic resonance structure where the vibrator is provided.

Specifically, in the example shown in fig. 10, the fan noise reduction system has a film type resonance structure 30a as an acoustic resonance structure, and a wind-proof member 48 disposed so as to cover the film 34 is provided on the surface of the diaphragm and the film 34 of the film type resonance structure 30 a.

The wind shielding member 48 is a member that passes sound and suppresses entry of wind. By disposing the wind-shielding member 48 on the surface of the film 34, it is possible to suppress the air flow generated by the fan from applying wind pressure to the film, which is the vibrator of the film-type resonance structure, to thereby shake the film, and to suppress the reduction of the noise reduction effect due to the wind.

As the wind-proof member 48, a foam such as a sponge, in particular, a porous structure such as a fiber body such as an open-cell foam, a cloth, or a nonwoven fabric can be used. In addition, as a rubber material film such as a silicone rubber film having an extremely small young's modulus, a thin plastic film having a thickness of about 10 μm such as a wrap film, or the like, a film that is loosely fixed without being strained and is characterized can be used. Since these membranes are quite different from the membrane 34 of the membrane type resonance structure in thickness, hardness and fixing method, there is no strong resonance in the audible range to pass sound.

In the example shown in fig. 1 to 3, the fan noise cancellation system is configured to have only the film-type resonance structure 30a, but the fan noise cancellation system is not limited to this, and may also have a porous sound absorbing material.

For example, a porous sound absorbing material may be provided in a space surrounded by the frame 32 and the film 34 of the film resonant structure 30a, that is, in the rear space 35. Alternatively, the film-type resonant structure 30a may have a porous sound absorbing material on the surface of the film 34.

By configuring the fan silencer system to have a porous sound absorbing material, it is possible to selectively damp sounds of frequencies other than the dominant sound of the resonator over a wide frequency band. Further, the porous sound absorbing material may be used as the wind shield member.

The porous sound absorbing material is not particularly limited, and a known porous sound absorbing material can be suitably used. For example, foaming materials such as foamed urethane, flexible urethane foam, wood, ceramic particle sintered material, phenol foam, and the like, and materials containing fine air; glass wool, rock wool, ultrafine fibers (shinsulite (tm) manufactured by 3 mcompray), floor mats, pile blankets, melt-blown nonwoven fabrics, metal nonwoven fabrics, polyester nonwoven fabrics, metal wool, felts, heat insulating boards, glass nonwoven fabrics, nonwoven fabric materials, kapok cement boards, silica nanofiber-based materials, gypsum boards, and other various known porous sound-absorbing materials.

The flow resistance of the porous sound-absorbing material is not particularly limited, but is preferably 1000 to 100000(Pa · s/m)²) More preferably 3000 to 80000(Pa · s/m)²) More preferably 5000 to 50000(Pa · s/m)²)。

The flow resistance of the porous sound-absorbing material can be evaluated by measuring the sound absorption rate at normal incidence of the porous sound-absorbing material having a thickness of 1cm and fitting the measured value to a Miki model (j.acoust.soc.jpn., 11(1)). pp.19-24 (1990)). Alternatively, the evaluation may be made in accordance with "ISO 9053".

Furthermore, a plurality of porous sound absorbing materials having different flow resistances can be laminated.

Here, in the example shown in fig. 1 to 3, the fan noise reduction system has the membrane type resonance structure 30a as the acoustic resonance structure, but the present invention is not limited to this. The fan muffling system may have a helmholtz resonance structure and/or an air column resonance structure as the acoustic resonance structure.

Fig. 11 is a schematic cross-sectional view showing an example of a fan silencer system having the structure of the helmholtz resonance structure 40. The fan muffler system shown in fig. 11 has the same configuration as the fan muffler system shown in fig. 4 except that a helmholtz resonance structure 40 is provided as an acoustic resonance structure instead of the film-type resonance structure 30 a.

In the example of fig. 11, the acoustic resonant structure is a helmholtz resonant structure 40. The helmholtz resonance structure 40 has: a frame 42 having a shape of a prism and having an opening portion with a bottom surface formed on one surface; and a plate-shaped cover 44 covering the opening surface of the frame 32 where the opening is formed, fixing the peripheral edge portion to the frame 32, and having a through hole 46. The helmholtz resonator 40 has the following structure: the air in the internal space 43 surrounded by the frame 42 and the lid 44 functions as a spring, and the air in the through-hole 46 formed in the lid 44 functions as mass, and performs resonance of the mass spring, and sound is absorbed by thermal viscous friction on the wall vicinity portion of the through-hole 46.

In the example shown in fig. 11, the cover 44 having the through-hole 46 is parallel to the rotation axis direction of the axial flow fan 12a, and the cover 44 is disposed toward the rotation axis side.

Conventionally, when a helmholtz resonance structure is used for sound attenuation, by matching the helmholtz resonance structure with the frequency of a sound to be attenuated, the sound of the frequency can be attenuated. Therefore, the noise cancellation effect for sounds in a frequency band other than the resonance frequency is low, and there is a problem that it is difficult to cancel a plurality of discrete frequency sounds generated by the fan.

In contrast, in the fan silencing system of the present invention, the helmholtz resonance structure 40 is disposed in the near field region of the sound generated by the fan, so that the above-described two interaction mechanisms can be generated to silence a plurality of discrete frequency sounds generated by the fan.

In the case where the helmholtz resonance structure 40 is used as the acoustic resonance structure, the resonance frequency of the helmholtz resonance preferably coincides with any one of the discrete-frequency sounds generated by the axial-flow fan 12 a.

The resonance frequency of the helmholtz resonance is determined by the volume of the internal space surrounded by the housing 42 and the cover 44, the area and length of the through-hole 46, and the like. Accordingly, the resonance frequency can be appropriately set by adjusting the volume of the internal space surrounded by the housing 42 and the cover 44 of the helmholtz resonance structure 40, the area, the length, and the like of the through hole 46.

Here, although the through-hole 46 is formed in the lid 44 in the example shown in fig. 11, the present invention is not limited to this, and the through-hole 46 may be formed in the housing 42. However, in this case, the inlet and outlet of the through-hole must be directed in the direction in which the discrete-frequency sound generated by the axial-flow fan 12a propagates, or in the flow path direction of the fan in fig. 11.

In the example shown in fig. 11, the helmholtz resonance structure 40 is a structure in which the housing 42 and the cover 44 are separate bodies, but the housing 42 and the cover 44 may be formed integrally.

In the helmholtz resonance structure 40, the air in the through-hole 46 is a vibrator, and the surface of the lid 44 having the through-hole 46 is a surface provided with the vibrator. Therefore, the surface of the lid portion 44 having the through-hole 46 is preferably arranged in parallel to an axis perpendicular to the air blowing port. Further, a wind-proof member may be disposed on the surface of the cover 44.

The shape of the helmholtz resonance structure 40 viewed in the direction perpendicular to the surface of the cover 44 may be a quadrangle, or a polygon such as a triangle, a circle, an ellipse, or the like.

In the example shown in fig. 11, the fan muffler system is configured to have two helmholtz resonance structures 40, but the fan muffler system is not limited to this configuration, and may have one helmholtz resonance structure, or may have 3 or more helmholtz resonance structures. In the case where the plurality of helmholtz resonance structures are provided, the housings of the respective helmholtz resonance structures may be integrally formed, or may be provided to share an internal space.

When the structure has a plurality of helmholtz resonance structures, the structure may have helmholtz resonance structures having different resonance frequencies.

Further, in the present invention, the resonator included in the muffler may be an air column resonance structure.

The air column resonance structure causes resonance by generating a standing wave in a resonance tube having an opening.

Conventionally, when an air column resonance structure is used for sound attenuation, sound of a frequency to be attenuated can be attenuated by matching the frequency of the air column resonance structure with the frequency of the sound. Therefore, the noise cancellation effect for sounds in a frequency band other than the resonance frequency is low, and there is a problem that it is difficult to cancel a plurality of discrete frequency sounds generated by the fan.

In contrast, in the fan noise cancellation system according to the present invention, the air column resonance structure is disposed in the near-field region of the sound generated by the fan, so that the two interaction mechanisms are generated to cancel the sound of a plurality of discrete frequencies generated by the fan.

In the case where the air column resonance structure is used as the acoustic resonance structure, the resonance frequency of the air column resonance is also preferably coincident with any one of the discrete frequency sounds generated by the fan.

The resonance frequency of the gas column resonance is determined by the length of the resonance tube, and the like. Therefore, the frequency of the resonance sound can be appropriately set by adjusting the depth of the resonance tube, the size of the opening, and the like.

In addition, when the acoustic resonance structure is configured to include an internal space and a through-hole (opening) that communicates the internal space with the outside, whether the acoustic resonance structure is a resonance structure that resonates with an air column or a resonance structure that resonates with helmholtz is determined by the size and position of the through-hole, the size of the internal space, and the like. Accordingly, by appropriately adjusting these, it is possible to select either one of the air column resonance and the helmholtz resonance.

In the case of the air column resonance structure, if the opening is narrow, the acoustic wave is reflected at the opening, and the acoustic wave is less likely to enter the internal space. Specifically, when the opening is rectangular, the length of the short side is preferably 1mm or more, more preferably 3mm or more, and still more preferably 5mm or more. When the opening is circular, the diameter is preferably within the above range.

On the other hand, in the case of helmholtz resonance, since it is necessary to generate thermal viscous friction in the through-hole, it is preferable to be somewhat narrow. Specifically, when the through-hole is rectangular, the short side length is preferably 0.5mm or more and 20mm, more preferably 1mm or more and 15mm or less, and further preferably 2mm or more and 10mm or less. In the case where the through-hole is circular, the diameter is preferably within the above range.

In addition, the fan silencer system of the present invention may be configured to have different types of acoustic resonant structures. For example, the structure may be configured to have a helmholtz resonance structure and a membrane type resonance structure.

Here, from the viewpoint of downsizing and thinning, it is preferable to use a membrane type resonance structure as the acoustic resonance structure.

Examples of the material of the frame body and the lid portion of the film-type resonance structure, the helmholtz resonance structure, and the air column resonance structure (hereinafter collectively referred to as "frame material") include a metal material, a resin material, a reinforced plastic material, and carbon fiber. Examples of the metal material include metal materials such as aluminum, titanium, magnesium, tungsten, iron, steel, chromium molybdenum, nickel chromium molybdenum, copper, and alloys thereof. Examples of the resin material include resin materials such as acrylic resin, polymethyl methacrylate, polycarbonate, polyamideimide, polyarylate, polyetherimide, polyacetal, polyetheretherketone, polyphenylene sulfide, polysulfone, polyethylene terephthalate, polybutylene terephthalate, polyimide, ABS resin (Acrylonitrile), Butadiene (Butadiene), Styrene (Styrene) copolymer synthetic resin), polypropylene, and triacetyl cellulose. Examples of the Reinforced plastic material include Carbon Fiber Reinforced Plastics (CFRP) and Glass Fiber Reinforced Plastics (GFRP). Further, there may be mentioned rubbers such as natural rubber, chloroprene rubber, butyl rubber, EPDM (ethylene/propylene/diene rubber), silicone rubber, and crosslinked structures containing these rubbers.

Various honeycomb core materials can be used as the frame material. Since the honeycomb core material is lightweight and used as a high rigidity material, the existing product is easily obtained. Honeycomb Core materials formed of various materials such as aluminum honeycomb Core, FRP honeycomb Core, paper honeycomb Core (Shin Nippon honeycomb Core co., Ltd, Showa air Industry co., Ltd, etc.), thermoplastic resin (PP, PET, PE, PC, etc.) honeycomb Core (GIFU input co., Ltd, teccelll, etc.) can be used as the frame body.

Further, as the frame material, a structure containing air, that is, a foam material, a hollow material, a porous material, or the like can be used. When a large number of resonators are used, the frame can be formed of, for example, a foam material containing independent cells so as not to allow air to flow between the cells. For example, various materials such as closed cell polyurethane, closed cell polystyrene, closed cell polypropylene, closed cell polyethylene, and closed cell rubber sponge can be selected. By using the independent cell body, sound, water, gas, and the like are not passed through and structural strength is large as compared with a continuous cell body, and therefore, the independent cell body is suitable for use as a frame material. When the porous sound-absorbing body has sufficient support, the frame body may be formed only of the porous sound-absorbing body, or materials such as those for the porous sound-absorbing body and the frame body may be used in combination, for example, by mixing or kneading. By using a material system containing air inside in this manner, the device can be made lightweight. Further, heat insulation can be provided.

Here, the frame material is preferably made of a material having higher heat resistance than the flame retardant material, from the viewpoint of being able to be disposed at a position where the temperature becomes high. The heat resistance can be defined by, for example, the time period in clause 108, clause 2, which satisfies the construction standards act regulations. The material according to item 108, item 2, which satisfies the construction Standard code of practical examples, is a flame retardant material when the time is 5 minutes or more and less than 10 minutes, a quasi-noncombustible material when the time is 10 minutes or more and less than 20 minutes, and a noncombustible material when the time is 20 minutes or more. However, heat resistance is generally defined for each field. Therefore, depending on the field in which the fan noise reduction system is used, the frame material may be made of a material having heat resistance equivalent to or higher than flame retardancy defined in the field.

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

As a material of the film 34, various metals such as aluminum, titanium, nickel, permalloy, 42 alloy, kovar, nickel-chromium, copper, beryllium, phosphor bronze, brass, nickel-silver, tin, zinc, iron, tantalum, niobium, molybdenum, zirconium, gold, silver, platinum, palladium, steel, tungsten, lead, and iridium; resin materials such as PET (polyethylene terephthalate), TAC (triacetylcellulose), PVDC (polyvinylidene chloride), PE (polyethylene), PVC (polyvinyl chloride), PMP (polymethylpentene), COP (cycloolefin polymer), ZEONOR, polycarbonate, PEN (polyethylene naphthalate), PP (polypropylene), PS (polystyrene), PAR (polyarylate), aramid, PPs (polyphenylene sulfide), PEs (polyethersulfone), nylon, PEs (polyester), COC (cyclic olefin copolymer), cellulose acetate butyrate, nitrocellulose, cellulose derivatives, polyamide, polyamideimide, POM (polyoxymethylene), PEI (polyetherimide), polyrotaxane (sliding ring material, etc.), and polyimide. In addition, glass materials such as film glass, and fiber-reinforced plastic materials such as CFRP (carbon fiber reinforced plastic) and GFRP (glass fiber reinforced plastic) can also be used. Further, rubbers such as natural rubber, chloroprene rubber, butyl rubber, EPDM, silicone rubber, and the like, and crosslinked structures containing these rubbers can be used. Alternatively, a combination of these materials may be used.

In the case of using a metal material, the surface may be plated with a metal in view of suppressing rust or the like.

From the viewpoint of excellent durability against heat, ultraviolet rays, external vibration, and the like, a metal material is preferably used as the material of the film 34 in applications requiring durability.

The method of fixing the film or the lid to the housing is not particularly limited, and a method using a double-sided tape or an adhesive, a mechanical fixing method such as screwing, or pressure bonding may be suitably used. The fixing method can be selected from the viewpoints of heat resistance, durability, and water resistance, as in the case of the frame material and the film. For example, as the adhesive, CEMEDINE co., LTD., "Super X" series, ThreeBond co., LTD., "3700 series (heat resistant)", TAIYO WIRE croth co., LTD manufactured heat resistant epoxy adhesive "Duralco series" and the like can be selected. Further, as the double-sided tape, high heat-resistant double-sided tape 9077 manufactured by 3M corporation and the like can be selected. In this way, various fixing methods can be selected for the required characteristics.

Here, although the fan noise reduction system has the axial flow fan 12a as a fan and suppresses noise of the axial flow fan (propeller fan) in the example shown in fig. 1 and the like, the fan noise reduction system is not limited to this, and can be applied to a conventionally known fan such as a sirocco fan, a turbo fan, a centrifugal fan, or a linear flow fan.

The sirocco fan receives air from the direction of the rotation axis of the rotor having blades and supplies the air in the direction perpendicular to the rotation axis, and has an air supply port on the side surface. Therefore, for example, as shown in fig. 12, when the fan is a sirocco fan 12b, the film-type resonance structure 30a (acoustic resonance structure) is disposed so as to be in contact with the air blowing port 38. The structure of the film-type resonance structure 30a is the same as that of the example shown in fig. 1 and the like.

In the example shown in fig. 12, the film-type resonance structure 30a is disposed at a position not to block the air blowing port of the sirocco fan 12 b. In the film-type resonance structure 30a, the film 34 is parallel to the direction perpendicular to the air blowing port of the sirocco fan 12b, and the film 34 is disposed toward the air blowing port.

In the case of the sirocco fan, since the sound is generated from the blade portion of the fan as described above, a region of a distance smaller than λ/4 from the blade portion of the fan is also a near-field region. Thus, by arranging the acoustic resonance structure in the near-field region, the above-described two interactions can be generated in the near-field region to obtain the noise reduction effect.

Examples

The present invention will be described in more detail below with reference to examples. The materials, the amounts used, the ratios, the contents of the treatments, the procedures of the treatments, and the like shown in the following examples can be appropriately changed without departing from the gist of the present invention. Accordingly, the scope of the present invention should not be construed as being limited by the examples shown below.

Comparative example 1

An axial fan (Model 109P0612K701 manufactured by SANYO DENKI CO., LTD.) was used as the fan. The axial flow fan has an outer diameter of 60mm × 60mm and a thickness of 15 mm. Since the fan is provided with the casing on the side of the exhaust direction, the distance from the end of the front surface of the air blowing port to the blades of the rotor is about 5 mm.

In order to suppress the influence of solid vibration from the fan, vibration-proof rubber having a thickness of 5mm was disposed below the fan. Further, in order to suppress sound emitted as solid vibration from the side of the fan, the side surface of the fan case was surrounded by acrylic having a thickness of 5 mm.

A square duct having a length of 30mm in the duct direction was produced by cutting and combining a rectangular plate having a short side of 30mm length with an acrylic plate having a thickness of 5mm to have an inner diameter of 60mm square equal to the outer diameter of the fan. The acrylic sheet was processed using a laser cutter.

The duct is disposed on the surface of the fan on the air outlet side so that the duct channel of the fan coincides with the cross section of the duct. The frame surrounding the fan case and the outside of the duct are connected and closed with an adhesive tape, thereby producing a structure in which the duct is in close contact with the fan as shown in fig. 13.

< determination >

The fan was driven using the fabricated structure and the volume was measured.

In the sound measurement, in a position separated by a distance of 200mm in the axial direction from the fan center position, in order to avoid the influence of wind, microphones are arranged at points deviated by 50mm from the center axis in the horizontal direction and the vertical direction (ACO co., ltd. manufactured 1/2inch microphone 4152). The microphones are disposed on both sides of the exhaust side and the air supply side.

The fan was driven using a dc regulated power supply. The fan driving conditions were set to 12V and 0.25A.

Fig. 14 shows the measurement results of the exhaust microphone. The horizontal axis of the graph shown in fig. 14 is a logarithmic display. As can be seen from fig. 14, the fan with the rotating blades is characterized by large peak sounds (narrow-band sounds) that occur at a plurality of frequencies. That is, it is known that discrete frequency sound is generated. Wherein, the large peak value is in integral multiple relation. In particular, the sound volumes of 1.1kHz and 2.2kHz are large.

The wind speed at the outlet-side end of the duct was measured by an anemometer, and the wind speed was 3.1 m/s. Hereinafter, until example 3, no change in wind speed was observed.

[ example 1]

A fan noise cancellation system was produced in the same manner as in comparative example 1, except that the inner wall of the duct was formed into a film-type resonance structure produced as follows. The resonance frequency of the film type resonance structure was set to 2.2 kHz.

< design of membrane type resonance structure >

A membrane-type resonance structure was designed by performing acoustic structure coupling calculation based on a finite element method using COMSOL multi-hydrogen ics (manufactured by COMSOL inc.). The film was designed with the material set to PET and the thickness set to 75 μm, and with the dimensions and back surface distance varied. It is found that in a film type resonance structure in which the inner diameter of a circular frame body, which is a vibration part of a film, is 24mm and the distance from the back surface is 6mm, the film type resonance structure has resonance at 2.2kHz and has high absorption.

It is understood that 6mm of the back surface distance corresponds to a distance of 0.038 × λ with respect to the wavelength λ of 2.2kHz, and resonance can be achieved with a very thin structure. In the case of a normal gas column resonance structure of a one-side closed tube, since the required length is 0.25 × λ, it is found that the thickness can be reduced to a size of about 15% with respect to the gas column resonance structure.

< manufacture of membrane type resonance structure >

The above-designed structure was manufactured by processing an acrylic plate with a laser cutting machine. Specifically, an acrylic plate having a thickness of 3mm was processed to produce two orifice plate members having an outer shape of a square of 30mm and having an opening portion of a diameter of 24mm therein and a square plate member having an outer shape of 30 mm. The two apertured plate members and the plate member were superposed in this order and bonded with a double-sided adhesive tape ("Power of the field" manufactured by ASKUL Corporation), thereby producing a frame body.

A PET film (Lumirror manufactured by TORAY INDUSTRIES, inc.) having a thickness of 75 μm was attached to the open surface of the frame body with a double-sided tape. A film type resonance structure having a square outer shape of 30mm square, an inner shape of 24mm, a thickness of 75 μm and a back surface distance of 6mm was produced by cutting a PET film in accordance with the outer shape of a frame.

6 of the film-type resonance structures were produced, and a tube (30 mm in length) in which 3 out of 4 surfaces of the tube were two film-type resonance structures was produced (see fig. 5).

< determination >

The fan of the fan noise reduction system thus fabricated 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.

Fig. 15 shows the exhaust side measurement results, and fig. 16 shows the intake side measurement results. Fig. 15 and 16 also show the results of comparative example 1.

As is clear from fig. 15, a large noise reduction effect of about 20dB can be obtained at a resonance frequency of 2.2kHz in the film type resonance structure. Further, it is understood that the noise cancellation effect can be obtained also for a plurality of discrete frequency sounds of different frequencies generated by the rotation of the fan as shown by arrows in fig. 15. That is, it is found that the noise cancellation effect can be obtained even at frequencies other than the resonance frequency of the film type resonance structure. As described above, it is understood that the fan noise canceling system according to the present invention can cancel the noise of the frequencies other than the resonance frequency of the acoustic resonance structure by disposing the acoustic resonance structure in the near-field region of the noise generated by the fan, and thus can cancel the noise of the plurality of discrete-frequency sounds of different frequencies generated by the rotation of the fan.

Further, it is found that the noise cancellation effect at the frequency can be further improved by matching the resonance frequency of the film type resonance structure with one of a plurality of discrete-frequency sounds of different frequencies generated by the rotation of the fan.

Further, as is clear from fig. 16, the sound volume is also reduced at the resonance frequency and other frequencies of the film type resonance structure on the intake side. That is, it is known that the silencing effect on the exhaust side does not reflect sound and output to the intake side, but silences the sound together with the exhaust side and the intake side. It is considered that this effect is produced by a phenomenon in which sound caused by membrane vibration is absorbed by the membrane type resonance structure and sound emission from the sound source is suppressed by interference of sound reflected according to the membrane type resonance structure with the sound source.

In the fan noise reduction system according to example 1, the distance from the sound source portion (blade) of the fan to the center of the membrane vibration portion of the membrane type resonance structure is "5 mm from the front surface of the blade of the fan to the front surface of the air blowing port" + "15 mm from the center position of the membrane type resonance structure to the front surface of the air blowing port of the fan" is 20 mm. Since the wavelength/4 with a frequency of 2.2kHz was 39mm, it was found that the film type resonance structure was disposed in the near field region.

Comparative example 2

As shown in fig. 17, comparative example 2 has the following structure: the film-type resonance structure 30a is disposed separately from the axial flow fan 12a, and the duct 100 is disposed between the film-type resonance structure 30a and the axial flow fan 12 a. The same structure as that of the film type resonance structure of example 1 is used for the film type resonance structure 30 a. The pipe 100 was the same as that of comparative example 1 except that the length was 60 mm.

In this structure, the distance from the sound source portion (blade) of the fan to the film type resonance structure was 80 mm. Accordingly, the film-type resonance structure 30a is configured to be disposed outside the near-field region.

< determination >

The volume was measured on the exhaust side and the intake side in the same manner as in comparative example 1 by driving the fan of the fan silencing system of comparative example 2. In comparative example 2, the sound volume measurement results when the membrane type resonant structure 30a was replaced with the duct were compared with each other, and the sound deadening volume was obtained from the difference.

The results are shown in fig. 18.

Fig. 19 shows the measurement results of the sound volume when the parts of the membrane type resonance structures 30a of comparative example 3 and comparative example 3 are replaced with pipes (simple pipes).

As is clear from fig. 18, in comparative example 2, the sound can be suppressed at the resonance frequency of the film-type resonance structure 30 a.

However, as is clear from fig. 19, which further expands the frequency range, in the structure of comparative example 3, the noise cancellation effect can be obtained at a frequency other than the resonance frequency of the film type resonance structure 30 a.

In comparative example 2, since the film type resonance structure and the sound source are separated by λ/2, the noise cancellation effect is exhibited by the interference effect (far-field interference) which is a normal sound fluctuation. On the other hand, since it is considered that the mechanism in the near field region is not generated, it is natural that it does not contribute to noise reduction other than the resonance frequency of the film type resonance structure.

In contrast, when the film type resonance structure is disposed in the near field region as in example 1, it is necessary to integrally deal with the interaction between the film type resonance structure and the sound source, and it is also necessary to consider the interaction of near field sound of high wave number that does not propagate to a far distance. In this case, it is considered that the above mechanism contributes to the release of sound at a frequency other than the resonance frequency of the membrane type resonance structure. Therefore, in the near field region, a sound deadening effect can be given to a wide band of sound.

From the above results, it is understood that, as in example 1 of the present invention, by disposing the film type resonance structure in the near field region, it is possible to mute a plurality of discrete frequency sounds generated by the fan. Further, it is found that a higher noise reduction effect can be obtained at one frequency of the discrete-frequency sound by matching the resonance frequency of the membrane type resonance structure with the frequency. Further, it is known that the fan noise can be suppressed without blocking the air passage.

[ example 2]

A study was made to change the peak sound frequency by changing the type of fan using the same membrane type resonance structure as in example 2. A DC axial fan "9 GA0612G 9001" (frame size 60mm, thickness 10mm) manufactured by SANYO DENKI co. The fan was fixed in the same manner as in example 1, and the case where the same film type resonance structure as in example 1 was attached to the exhaust side (example 2) and the case where a duct having the same duct length and a length of 30mm was attached at the same position instead of the resonance structure (comparative example 3) were measured, respectively.

The measurement results are shown in fig. 20. In the case of this fan, the frequency of the peak sound appears at a frequency deviated from the resonance frequency of the film type resonance structure. In the vicinity of the resonant frequency of 2.2kHz of the film type resonant structure, the noise reduction of about 8dB appears widely. On the other hand, it is found that at the peak sound frequency of the fan (1.2kHz, 2.4kHz, 3.6kHz), when each film type resonance structure is in the near field region, the sound can be suppressed by the original peak sound volume.

In this way, it is understood that the peak sound of the fan having a frequency different from the resonance frequency of the film type resonance structure can be suppressed by the resonance structure in the near field region.

In addition, as for the peak sound canceling volume, it is found that the canceling volume in the case of example 1 in which the resonance frequency is made to coincide with the fan peak sound frequency is preferably larger than that in the case of the present example in which the resonance frequency is shifted from the fan peak sound frequency.

[ example 3]

A film type resonance structure was produced in the same manner as in example 1, except that the resonance frequency of the film type resonance structure was set to 1.1 kHz.

< manufacture of membrane type resonance structure >

As a result of designing by the finite element method using COMSOL multi-resonance, it was found that the resonance frequency was 1.1kHz by setting the back surface distance of the film-type resonance structure of example 1 to 6mm to 15 mm. An acrylic plate was processed by a laser cutter, and a film type resonance structure was produced in the same manner as in example 1.

The produced film-type resonance structure was disposed at a position 30mm apart from the surface of the air outlet of the fan. A pipe (duct) is connected between the film type resonance structure and the fan (refer to fig. 6). The distance from the center of the film type resonance structure to the sound source portion (blade) of the fan was 50 mm. On the other hand, since the wavelength/4 with the frequency of 1.1kHz was 78mm, it was found that the film type resonance structure was disposed in the near field region.

< determination >

The fans of the manufactured fan noise reduction systems were driven, and the sound volumes were measured on the exhaust side and the intake side in the same manner as in example 1.

The results are shown in fig. 21. Fig. 21 also shows the measurement results of the sound volume when the membrane resonance structure of example 3 is replaced with a duct (simple duct).

As is clear from fig. 21, a large noise reduction effect of about 10dB can be obtained at a resonance frequency of 1.1kHz in the film resonance structure. Further, it is known that a sound deadening effect can be obtained also for a plurality of discrete frequency sounds generated by the fan.

[ example 4]

A fan noise cancellation system was fabricated in a similar manner to that of example 1, except that the film type resonance structure fabricated in example 3 was disposed downstream of the film type resonance structure of the fan noise cancellation system of example 1 (see fig. 8).

The results are shown in fig. 22. Fig. 22 also shows the measurement results of the sound volume when the membrane resonance structure of example 4 is replaced with a duct (simple duct).

It is found that even at the resonance frequencies of 1.1kHz and 2.2kHz of the film type resonance structure, a large noise reduction effect of about 15dB can be obtained. That is, even when the membrane type resonance structures are arranged in series, the respective noise cancellation effects are exhibited.

It is also understood that the noise cancellation effect can be obtained even for a plurality of discrete frequency sounds generated by the fan shown by the arrow in fig. 22. That is, it is found that the noise cancellation effect can be obtained even at frequencies other than the resonance frequency of the film type resonance structure.

The difference between the two data of fig. 22 is obtained and shown in fig. 23 as the noise reduction amount. It is found that the noise peak of the fan is suppressed by 15dB or more in the vicinity of 1.1kHz and 2.2kHz, and that the noise reduction effect can be obtained also in other frequency bands.

With respect to the fan noise cancellation system of embodiment 4, in order to evaluate the magnitude of noise heard through the ears, octave band evaluation and overall noise amount evaluation are shown. In fig. 24, evaluation is performed every 1/3-fold frequency band, and the result of a characteristic evaluation (unit is dBA) in which the sound volume is set to correction in consideration of the sensitivity of the human ear is shown. By canceling the noise peaks at 1.1kHz, 2.2kHz and other frequencies, it is found that the overall sound is reduced even in the 1/3-fold frequency band evaluation in which the frequencies are widely averaged and evaluated. Then, the noise level is calculated by integrating the a characteristic correction for the entire frequency audible range. A noise level of 81.9(dBA) in the case of a simple pipe can reduce the noise level to 74.9(dBA) in the fan silencer system of embodiment 4. Since a noise level difference of 3dBA is sufficiently noticeable by a general person, the silencing effect of 7dBA is a level sufficiently quiet in the sense of body.

The following is thus shown: by conducting research for suppressing the discrete frequency sound generated from the fan and disposing the acoustic resonance structure in the near-field region, it is possible to suppress not only the resonance frequency sound but also the entire discrete frequency sound generated from the fan, thereby obtaining a large noise reduction effect.

[ example 5]

In comparison with examples 1 to 4, the type of the fan was changed to perform the measurement under the strong wind condition. A 9GA0612P1J03 (38 mm thick) fan manufactured by SANYO DENKI co. Fig. 25 shows the wind speed when the amount of current supplied to the fan is changed. By increasing the amount of current, a high wind speed and a high wind volume can be obtained.

A film type resonance structure having the same structure as that of example 2 was disposed on the exhaust side of the fan. However, the film surface of the film-type resonance structure was reduced by 5mm on the outer peripheral side of example 2 (see fig. 26). This is to dispose the wind-proof member in embodiment 6 later.

< determination >

The fan of the fan noise reduction system thus fabricated 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 fig. 27. Further, as comparative example 4, the measurement results obtained when the membrane-type resonance structure of example 5 was used instead of the pipe are also shown. In example 5 and comparative example 4, the structural lengths in the flow path direction were all 30mm and equal.

Then, the wind speed at the outlet side end of example 4 and comparative example 4 was measured using a wind speed meter. As a result, it was confirmed that the wind speed was 14.5m/s, and the wind speed was not changed in the case of mounting the film type resonance structure and the drum structure.

As is apparent from fig. 27, as indicated by arrows in fig. 27, the noise cancellation effect can be obtained at the peak of the frequency other than the resonance frequency of the film type resonance structure. However, it is known that the peak around the resonance frequency of 1.1kHz has an effect of amplifying sound at frequencies around the peak, and that the peak noise reduction effect is hardly obtained.

In example 5, the fan was rotated with a large air volume, and therefore the air became unstable. The wind applies wind pressure to the membrane surface, and generates vibration caused by the wind on the membrane surface. The vibration occurring on the membrane includes a wide frequency spectrum, but a resonance phenomenon occurs in a frequency in which resonance is designed in the design of the membrane type resonance structure, that is, a frequency intended to be silenced and the periphery thereof. At this resonance frequency, vibration generated on the film surface tends to remain for a long time, and the amplitude tends to be amplified even in a state where the fan is continuously operated. From here, sound is thus transmitted as a loudspeaker. In this way, it is considered that when a strong air volume is generated in the vicinity of the fan, the sound is amplified near the resonance frequency, and the target noise reduction effect is hardly obtained.

[ example 6]

A fan noise cancellation system was fabricated in the same manner as in example 5, except that a wind-proof member was disposed on the surface of the film type resonance structure in the fan noise cancellation system of example 5 (see fig. 10).

Urethane sponge (5 mm in thickness) was used as the wind-proof member. In order to prevent the influence on the membrane vibration as much as possible, a double-sided tape or the like is not used on the surface of the sponge on the membrane side, and a transparent tape is used to attach a part of the air side surface of the sponge (corresponding to the position of the frame portion of the membrane type resonance structure at the lower portion of the sponge) to the side wall portion of the membrane type resonance structure so that the sponge does not deviate from the membrane type resonance structure.

< determination >

The fan of the fan noise reduction system thus fabricated 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 fig. 28. The measurement results of comparative example 4 are also shown.

Then, the wind speed at the outlet side end of example 6 was measured using a wind speed meter. As a result, it was confirmed that the wind speed was 14.5m/s and the wind speed was unchanged.

As is clear from fig. 28, amplification of the sound near the resonance frequency (1.1kHz) generated in example 5 can be greatly suppressed. Further, as shown by arrows in fig. 28, it is understood that the effect of reducing peak sounds at frequencies other than the resonance frequency can be obtained. In fig. 28, it is understood that noise can be suppressed in a wide frequency band in a high frequency region of 5.4kHz or more. This is a sound absorption effect by the sponge disposed on the film surface.

From the above results, it is understood that when the membrane type resonance structure is disposed in the vicinity of the fan by disposing the wind shielding member on the membrane surface, the phenomenon of causing the sound to be emitted in the vicinity of the resonance frequency can be greatly suppressed. Further, it is found that the sound deadening effect of the porous sound absorbing material and the sound deadening effect by the film resonance structure can be simultaneously achieved by using the porous sound absorbing material as the wind shield member.

[ example 7]

A fan muffler system was produced in the same manner as in example 5, except that a helmholtz resonance structure was used as the acoustic resonance structure.

As a result of designing a Helmholtz resonance structure having a resonance frequency of 1.1kHz, the length of the through-hole was 3mm, the diameter of the through-hole was 4mm, the thickness of the inner space was 12mm, and the diameter of the inner space was 24 mm.

In order to achieve such a structure, a helmholtz resonance structure is manufactured by machining an acrylic plate with a laser cutter. A fan noise reduction system was produced in the same manner as in example 5 such that 6 units of the helmholtz resonance structure constituted the wall surface of the duct.

Fig. 29 shows the measurement result when the current amount supplied to the fan was 0.3A. Further, the measurement results obtained when the same length of pipe was installed instead of the helmholtz resonance structure are shown (comparative example 5). At this time, the wind speed was 5.5 m/s.

As is apparent from fig. 29, even when the helmholtz resonance structure is used as the acoustic resonance structure, a noise cancellation effect can be obtained for peak sounds at frequencies other than the frequency. On the other hand, the noise cancellation amount is small for the peak of the resonance frequency 1.1kHz, and sound amplification occurs in the periphery thereof. This is an effect of causing resonance in wind noise generated in the through hole of the helmholtz resonance structure at the resonance frequency of the resonance structure, thereby amplifying sound and sounding sound.

[ example 8]

The sound volume was measured in the same manner as in example 7, except that the amount of current supplied to the fan was set to 1.3A. The measurement results are shown in fig. 30. Further, the measurement results when the same length of pipe was installed instead of the helmholtz resonance structure are also shown (comparative example 6). The wind speed was 15.1 m/s.

As is clear from fig. 30, the effect of canceling a plurality of peak sounds of frequencies other than the resonance frequency can be obtained even in the helmholtz resonance structure at a high air volume. On the other hand, it is known that the wind noise whose resonance is amplified becomes larger by the high wind speed, and the peak sound in the vicinity of the resonance frequency is amplified.

As is apparent from the above description, the effect of being able to cancel sound of a plurality of discrete frequencies by the resonance structure is not limited to the film resonator, but is common. Further, it is considered that the amplification effect by the wind noise of the helmholtz resonance is larger than the phenomenon of the membrane type resonance structure ringing, and therefore, the membrane type resonance structure is more preferable particularly in the case of use in strong wind.

Comparative example 7

In order to study the application to fans other than axial flow fans, the application to a sirocco fan for a blower is studied. A blower 9BMC12P2G001 manufactured by SANYO DENKI co. The blower fan was disposed on a vibration-proof rubber having a thickness of 10mm, and was configured to discharge air sucked from above in a horizontal direction. An acrylic plate having an opening of the same size as the air blowing port (an opening of about 30mm × 52 mm) and a thickness of 5mm was disposed as a vertical baffle 102 at a position 30mm apart from the air blowing port, and the measurement microphone MP was disposed in such a manner that wind did not directly blow to the tip thereof, and experiments were carried out. The wind speed measured at the opening of the vertical baffle 102 at this time was 7.7 m/s.

The measurement was performed in a state where the air blowing port and the opening of the vertical baffle 102 were connected by a duct 100 made of an acrylic plate having a thickness of 5 mm. A schematic diagram is shown in fig. 31.

[ example 9]

A fan silencer system was fabricated in the same manner as in comparative example 7, except that 4 film-type resonance structures 30a of example 4 were arranged in a duct-like manner between the air blowing ports and the opening of the vertical baffle 102 (see fig. 32).

The minimum distance between the film-type resonance structure 30a and the blades of the sirocco fan is 24mm, and the film-type resonance structure 30a is disposed in the near field region.

< determination >

In example 9 and comparative example 7, the volume was measured by the measuring microphone MP while driving the fan.

The measurement results are shown in fig. 33.

As is clear from the results shown in fig. 33, in the structure of example 9, the peak sound can be reduced in the vicinity of the resonance frequency, and the silencing effect is exhibited also for peak sounds appearing at other frequencies. As a result, it was found that, even with the sirocco fan, the sound deadening effect of the sound of a plurality of discrete frequencies can be obtained by disposing the acoustic resonance structure in the near field region, as in the case of the axial flow fan.

From the above results, the effect of the present invention is remarkable.

Description of the symbols

10-fan silencer system, 12 a-axial fan, 12 b-multiblade fan, 16-housing, 16 a-blower, 18-rotor, 20-shaft, 22-blade, 26-duct, 30a, 30 b-membrane type resonance structure, 32, 42-frame, 34-membrane, 35-back space, 36-vibration-proof member, 38-blower, 40-helmholtz resonator, 43-interior space, 44-cover, 46-through hole, 48-wind-proof member, 100-duct, 102-vertical baffle, MP-microphone.