Multi-blade noise reduction cross-flow fan blade

Abstract

The utility model discloses a multi-blade noise reduction cross-flow fan, which belongs to the technical field of centrifugal fans and comprises a hub and a central shaft connected with the hub through a rubber shaft sleeve. Blades are evenly distributed in the circumferential direction of the hub, and the number of the blades is large. The blade profile of the blade is an asymmetric airfoil profile; the roots of the blades are connected with the hub in a welded mode, and micropore noise reduction structures are arranged on the surfaces of the blades. By means of the mode, the 37 asymmetric wing-shaped blades are evenly distributed in the circumferential direction, the front edge fillet and rear edge curvature optimization design is combined, blade surface airflow separation is effectively restrained, the vortex intensity and pressure pulsation in a wake area are reduced, and rotation noise caused by periodic excitation of the blades and wake flow non-uniformity is reduced fundamentally. The micropore noise reduction structures distributed on the surface of the blade in a gradient mode restrain turbulence development by disturbing the boundary layer and are matched with the inner side arc-shaped flow guide grooves to guide airflow to flow smoothly, and coupling noise of blade top leakage flow and boundary layer separation vortex is further weakened.

Description

Technical Field

[0001] This utility model relates to the field of centrifugal fan technology, specifically to a multi-blade noise-reducing cross-flow fan blade. Background Technology

[0002] The rotational noise of a cross-flow fan, also known as discrete frequency noise or blade passage frequency noise, is mainly caused by the following two factors:

[0003] Firstly, when the impeller rotates in free space, the blades interact with the surrounding gas medium, constantly impacting the gas. The air at a fixed position near the blades is periodically excited by the blades and their pressure field, thus generating noise. This is the essence of rotational noise.

[0004] Secondly, blade wakes also contribute to rotational noise. As airflow passes over the blades, a boundary layer forms on the blade surface, particularly thickening at the suction edge and generating numerous eddies. At the blade trailing edge, the boundary layers from the suction and pressure edges merge, forming a wake region. Within this region, the airflow pressure and velocity are significantly lower than in the main airflow region. During impeller rotation, the airflow in the blade exit region exhibits considerable non-uniformity. This non-uniform airflow periodically acts on the surrounding medium, causing pressure pulsations and thus generating noise.

[0005] In addition, if the distance between the moving blade row and the volute tongue is too small, the wake of the front blade row sweeping over the rear blade row, or the wake of the moving blade row sweeping over the volute tongue, will generate pulsating force, which will then cause rotational noise.

[0006] Based on this, this utility model designs a multi-blade noise-reducing cross-flow fan blade to solve the above problems. Utility Model Content

[0007] To address the aforementioned shortcomings of existing technologies, this utility model provides a multi-blade noise-reducing cross-flow fan blade, comprising:

[0008] The hub and the central shaft are connected to the hub via a rubber bushing.

[0009] The hub has 37 blades evenly distributed around its circumference.

[0010] The blade has an asymmetric airfoil shape, and the ratio of the radius of curvature R1 at the leading edge to the radius of curvature R2 at the trailing edge is 1:1.2-1.5.

[0011] The root of the blade is connected to the hub by welding, and the blade surface is provided with a microporous noise reduction structure, which is distributed in a gradient along the chord length of the blade.

[0012] Furthermore, the blade is made of glass fiber reinforced nylon composite material (such as glass fiber reinforced PA66+25-35%GF), and the inner side of the blade is provided with an arc-shaped guide groove.

[0013] Furthermore, the contact surface between the rubber bushing and the central shaft is provided with a spiral groove.

[0014] Furthermore, the asymmetric airfoil has a rounded leading edge.

[0015] Furthermore, the gradient density of the microporous noise reduction structure varies with the radial position of the blade.

[0016] To achieve the above objectives, this utility model provides the following technical solution:

[0017] Compared with the prior art, the advantages of this utility model are as follows:

[0018] 1. This invention employs 37 asymmetric airfoil blades evenly distributed circumferentially, combined with a leading-edge rounded corner and trailing-edge curvature optimization design, effectively suppressing airflow separation on the blade surface, reducing vortex intensity and pressure pulsation in the wake region, and fundamentally reducing rotational noise caused by periodic blade excitation and wake inhomogeneity. The gradient-distributed microporous noise reduction structure on the blade surface suppresses turbulence development by perturbing the boundary layer, and, in conjunction with the inner arc-shaped guide groove, guides the airflow smoothly, further weakening the coupling noise between the tip leakage flow and the boundary layer separation vortex.

[0019] 2. The central shaft and hub of this utility model are connected by a rubber bushing with a spiral groove. The viscoelastic damping of the rubber material and the frictional damping of the groove structure work together to effectively attenuate axial and circumferential vibration energy, reduce vibration transmission rate, and avoid resonance problems caused by impeller imbalance or aerodynamic load fluctuations. Attached Figure Description

[0020] To more clearly illustrate the technical solutions in the embodiments of this utility model or the prior art, the drawings used in the description of the embodiments or the prior art will be briefly introduced below. Obviously, the drawings described below are only some embodiments of this utility model. For those skilled in the art, other drawings can be obtained based on these drawings without creative effort.

[0021] Figure 1 This is a perspective view of a multi-blade noise-reducing cross-flow fan blade according to the present invention.

[0022] Figure 2 This is a schematic diagram of the tensile test results of the multi-blade noise reduction crossflow fan blade of this utility model;

[0023] Figure 3 This is a test diagram of the tensile force curve of the multi-blade noise-reducing cross-flow fan blade of this utility model;

[0024] Figure 4 This is a physical image of the multi-blade noise-reducing cross-flow fan blade of this utility model;

[0025] Figure 5 This is a schematic diagram of the blade structure of the multi-blade noise-reducing cross-flow fan blade of this utility model.

[0026] Figure 6 This is a noise test spectrum of the multi-blade noise-reducing cross-flow fan blade of this utility model.

[0027] The labels in the diagram represent:

[0028] 1. Hub; 2. Central shaft; 3. Blade; 4. Rubber bushing; 5. Arc-shaped guide groove. Detailed Implementation

[0029] To make the objectives, technical solutions, and advantages of the embodiments of this utility model clearer, the technical solutions of the embodiments of this utility model will be clearly and completely described below with reference to the accompanying drawings. Obviously, the described embodiments are only some, not all, of the embodiments of this utility model. All other embodiments obtained by those skilled in the art based on the embodiments of this utility model without creative effort are within the scope of protection of this utility model.

[0030] In some embodiments, please refer to the accompanying drawings. Figures 1-6 A multi-blade noise-reducing cross-flow fan blade includes: a hub 1 and a central shaft 2, wherein the central shaft 2 is connected to the hub 1 through a rubber bushing 4, and the rubber bushing has a Shore hardness of 70-80HA.

[0031] The damping ratio of rubber materials first increases and then decreases with increasing hardness. At 70-80HA, the ζ value is in the plateau region (about 0.15-0.20), which combines elastic deformation ability (absorbing vibration energy) and rapid recovery ability (avoiding continuous deformation).

[0032] When the hardness is <70HA, the rubber is too soft (e.g., 60HA). Although the damping ratio is slightly higher (ζ=0.22), the axial compression deformation is >1.5mm (under rated load), resulting in a wheel hub coaxiality deviation >0.3mm, which causes additional vibration and increases the risk of resonance.

[0033] When the hardness is greater than 80HA, the rubber is too hard (e.g., 90HA), the damping ratio drops sharply to below 0.10, the vibration energy is difficult to dissipate, the vibration transmission rate between the central shaft and the hub is greater than 70% (the transmission rate is ≤50% when the hardness is 70HA), and the vibration reduction effect is lost.

[0034] When the rubber bushing is rotating at high speed, it bears alternating centrifugal loads. The dynamic elastic modulus corresponding to the hardness of 70-80HA keeps the natural frequency of the bushing away from the impeller excitation frequency (e.g., the passing frequency of a 37-blade impeller = 37 × rotational speed / 60, which is about 86Hz at 1400rpm), thus avoiding resonance.

[0035] The rubber bushing and the hub mounting groove have an interference fit of 0.2-0.3mm. The rubber with a hardness of 70-80HA can fill the fit gap through elastic deformation, while generating sufficient contact pressure (such as 0.5-0.8MPa) to prevent axial movement (such as critical axial force ≥500N).

[0036] 37 blades 3 are evenly distributed around the hub 1. The blades 3 have an asymmetric airfoil shape, and the ratio of the leading edge radius of curvature R1 to the trailing edge radius of curvature R2 is 1:1.2-1.5.

[0037] According to the Hiemens stagnation flow theory, the flow near the leading edge can be regarded as a two-dimensional stagnation flow, and its critical radius of curvature for boundary layer separation satisfies:

[0038] in:

[0039] ν is the kinematic viscosity of air (1.5 × 10⁻⁶). −5 m 2 / s);

[0040] U ∞ The incoming flow velocity (takes a common engineering value of 20-40 m / s);

[0041] α cr The critical separation angle (experimental statistical value 8-12°);

[0042] Substituting the values, we can see that the lower limit of R1 is determined by the flow separation conditions. Considering the engineering safety margin, we initially determine that R1 > 0.8 mm (corresponding to R2 > 0.96 mm, i.e., R1:R2 > 1:1.2).

[0043] According to wake dynamics, the trailing vortex shedding frequency f and the trailing radius of curvature R2 satisfy the Strauhal number relationship: To avoid the vortex shedding frequency and the blade passage frequency Resonance requires The derivation yields R2 < 1.5R1 (corresponding to R1:R2 < 1:1.5).

[0044] A parametric airfoil is adopted, with a fixed chord length L=80mm. R1=1mm is set as the baseline, and R2 varies in increments of 1.2-1.5mm (covering a 1:1.2-1.5 ratio). The maximum airfoil thickness is fixed at 30%-35% of the chord length. The leading / trailing edge regions have a finer mesh (minimum mesh size 0.1mm). The overall mesh count is 150,000-200,000, satisfying the y... + <50 (applicable to wall functions), the following parameters are extracted through simulation for comparison:

[0045]

[0046] C L ×(1 / C D The maximum value appears in the range of 1:1.2-1.5 (5%-8% higher than the boundary value). The peak value of the wake vorticity decreases by 20%-30% in this range, indicating that the vortex shedding noise potential is the lowest and the boundary layer separation state transitions from no leading edge separation to controllable trailing edge separation.

[0047] The root of the blade 3 is connected to the hub 1 by laser welding, and the axial tensile strength at the weld is ≥5000N, as shown in the following data:

[0048]

[0049] The surface of blade 3 is provided with a microporous noise reduction structure, which is distributed in a gradient along the chord length of the blade, with a gradient density of 3-5 pores per centimeter.

[0050] The inner side of blade 3 is provided with an arc-shaped guide groove 5. The depth of the guide groove is 1 / 10 to 1 / 8 of the blade thickness. The core purpose of setting the arc-shaped guide groove 5 on the inner side of blade 3 is to guide the fluid (gas / liquid) to flow smoothly along the blade surface and avoid flow separation caused by the thickening of the boundary layer. According to the boundary layer theory, when the fluid flows through the curved surface of the blade, the near-wall velocity will gradually decrease due to viscosity, forming a velocity gradient layer (boundary layer). If the depth of the arc-shaped guide groove 5 is insufficient (such as less than 1 / 10 of the blade thickness), the disturbance and guidance effect of the groove on the boundary layer will be weak, and it will not be able to effectively change the local flow field distribution. It is easy to form vortices at the tail of the blade, increasing drag (such as pressure drag or induced drag). If the depth is too large (such as more than 1 / 8), the fluid will form local turbulence or secondary flow in the arc-shaped guide groove 5, which will aggravate energy dissipation.

[0051] The depth of the arc-shaped guide channel 5 directly affects the pressure distribution of the fluid within the channel. According to Bernoulli's equation, as the fluid flows through the arc-shaped guide channel 5, changes in the channel cross-section lead to a conversion between velocity and pressure: if the depth is too shallow, the velocity increase within the arc-shaped guide channel 5 is insufficient, resulting in limited pressure regulation; if the depth is too deep, the velocity within the arc-shaped guide channel 5 becomes excessively high, creating a low-pressure zone, and may even induce cavitation (in liquid media) or cavitation phenomena, damaging the surface of the blade 3. A depth range of 1 / 10 to 1 / 8 ensures that the fluid velocity gradient within the channel is within a reasonable range, achieving both the correction of the mainstream flow and avoiding abnormal pressure fluctuations.

[0052] The contact surface between the rubber bushing 4 and the central shaft 2 is provided with a spiral groove (not shown in the figure), with a groove depth of 0.3-0.5mm, to enhance the damping and shock absorption effect. The shock absorption effect of the rubber bushing mainly comes from the viscoelastic damping of the material. The spiral groove enhances the damping effect in the following ways:

[0053] The spiral surface of the groove causes the rubber to undergo three-dimensional composite deformation (tension, bending, and torsion) when under pressure or shear. Compared with planar contact, the deformation path is longer and more complex, which intensifies the internal friction between rubber molecular chains and improves the efficiency of energy conversion into heat energy (increases the damping loss factor tanδ).

[0054] When the central shaft vibrates, the helical edge of the groove will produce local micro-slippage with the shaft surface, which will supplement the viscoelastic damping of the material itself through frictional damping (Coulomb damping).

[0055] The spiral structure can change the propagation direction of vibration waves in the bushing, causing the axial vibration energy to diffuse circumferentially. It increases energy dissipation through wave scattering and mode conversion (such as longitudinal wave to shear wave). When the groove depth is less than 0.3 mm, the structure has a weak modulation effect on wave propagation; if it exceeds 0.5 mm, the stiffness may drop sharply due to the excessively thin rubber matrix, which will reduce the damping stability.

[0056] The effective damping range of rubber is positively correlated with its deformation, but it is necessary to avoid entering a state of plastic deformation or failure.

[0057] The lower limit is 0.3mm to ensure that the groove undergoes effective elastic deformation (strain ≥5%, estimated based on the Shore A hardness of rubber 70A, corresponding to a stress of about 0.3-0.5MPa) under the rated load, so that the molecular chain segment movement is fully activated and significant viscoelastic loss is generated.

[0058] The upper limit is 0.5mm to prevent the rubber thickness at the bottom of the groove from being too thin (assuming the bushing base thickness is ≥2mm, the remaining thickness is ≥1.5mm), and to avoid tearing failure caused by stress concentration (based on the rubber elongation at break being ≥300%, a safety margin of ≥60% is maintained).

[0059] Based on a simplified model of a spring-damped system, the relationship between the groove depth h and the damping coefficient c can be established using the following dimensionless parameter:

[0060]

[0061] The spiral groove increases the contact area (circumferential unfolded length × number of grooves × depth), thereby increasing the shear stiffness and damping coefficient of the rubber. Engineering experience shows that when h / t∈[0.15,0.25]

[0062] (Assuming the bushing wall thickness is 2-3 mm, corresponding to h = 0.3-0.75 mm), the damping performance increases linearly with depth. After exceeding 0.25 mm, the damping saturates due to the decrease in stiffness. Therefore, the intermediate high-efficiency range of 0.3-0.5 mm is selected.

[0063] The asymmetric airfoil features a rounded leading edge (not shown in the figure). A sharp leading edge, at small angles of attack or low-speed flow, is prone to premature boundary layer separation due to abrupt curvature changes, forming leading-edge vortices and causing a surge in pressure drag. The rounded corners, by smoothing the curvature transition (transforming the sharp leading-edge corner into a rounded surface), reduce the velocity gradient of the airflow around the leading edge, delay the boundary layer separation point, and make the pressure distribution near the leading edge more uniform, reducing the leading-edge suction peak and thus lowering the frictional drag from the local high-speed region.

[0064] Furthermore, the gradient density of the microporous noise reduction structure varies with the radial position of the blade, with 3 pores per centimeter in the leaf root region and 5 pores per centimeter in the leaf tip region.

[0065] Experimental Example 1: Noise Detection of Multi-blade Noise Reduction Crossflow Fan;

[0066] The test was conducted in a semi-anechoic chamber conforming to ISO 3744 standards, with background noise ≤20dB(A).

[0067] Temperature: 25±2℃, Humidity: 50±5%RH.

[0068] Test equipment:

[0069] High-precision sound level meter (Level 1 accuracy, frequency range 20Hz-20kHz).

[0070] Spectrum analyzer (including 1 / 3 octave band analysis function);

[0071] Anemometer (range 0-20m / s, accuracy ±0.1m / s);

[0072] Tachometer (used to monitor the speed of fan blades).

[0073] Install the fan blade under test inside the standard cross-flow fan housing, ensuring that the coaxiality error between the central shaft and the motor connection is ≤0.1mm, the gap between the fan blade and the housing is uniform to avoid airflow leakage, the test speed is 2000rpm (rated operating conditions), and the air volume setting is adjusted to the target value through the frequency converter;

[0074] Set up the main measuring point 1m directly in front of the fan outlet;

[0075] Four auxiliary measuring points are set every 1m around the circumference of the wind turbine, with the height aligned with the center of the wind turbine.

[0076] Data collection steps:

[0077] Start the fan and allow it to run stably for 5 minutes;

[0078] The A-weighted sound pressure level (dB(A)) at each measuring point was measured using a sound level meter. Each measurement lasted 30 seconds, and the average value was taken.

[0079] Comparative experimental design:

[0080] (1) 37 blades + asymmetric airfoil + microporous noise reduction + no guide groove;

[0081] (2) 37 blades + asymmetric airfoil + microporous noise reduction + flow guide groove;

[0082] (3) 37 blades + asymmetric airfoil + no microporous noise reduction + guide channel;

[0083] (4) 35 blades + asymmetric airfoil + microporous noise reduction + flow channel.

[0084] The results are as follows Figure 6 As shown, changing the number of blades from 35 to 37, as well as the addition of micro-holes and arc-shaped guide grooves, significantly reduced noise in the wind turbine blades.

[0085] Increasing the number of blades reduces the spacing between adjacent blades, resulting in a more even distribution of airflow as it passes through the blades. This prevents uneven distribution of airflow velocity, avoiding localized areas of excessively high or low velocity. A more uniform airflow distribution reduces the pressure gradient on the blade surface, decreasing the tendency for boundary layer separation and thus suppressing vortex formation and shedding. This reduces vortex shedding noise sources. At the same rotational speed, the blades cut the airflow at a higher frequency, resulting in a more even distribution of periodic disturbance forces on the airflow. This helps reduce the amplitude of airflow pulsations, lowers broadband noise caused by airflow pressure fluctuations, and makes the blades more stable under stress, reducing structural noise caused by blade vibration.

[0086] When airflow passes through the micropores on the blade surface, the geometric constraint of the micropores causes strong local disturbances and vortices in the airflow. The formation and rupture of these micro vortices convert the macroscopic kinetic energy of the airflow into other forms of energy such as microscopic thermal and acoustic energy, and dissipates the energy through heat conduction and viscous dissipation through the orifice walls, thereby reducing the intensity and scale of vortex shedding at the blade trailing edge and reducing the noise it generates.

[0087] The micropores are distributed in a gradient along the blade chord, which can cause varying degrees of disturbance to the boundary layer development of the airflow at different locations on the blade surface. In the leading edge region of the blade, the higher density of micropores can induce a slight transition in the boundary layer earlier, allowing the airflow to adapt to the flow state on the blade surface sooner. In the trailing edge region of the blade, the gradually decreasing density of micropores helps to stabilize the boundary layer, delay airflow separation, optimize the airflow path on the blade surface, and reduce eddies and noise caused by boundary layer separation.

[0088] The specific shape and size of the arc-shaped guide groove (depth of 1 / 10-1 / 8 of the blade thickness) can form a streamlined channel on the inner side of the blade that is conducive to airflow adhesion. When the airflow passes through the guide groove, its flow direction and velocity are guided, causing the airflow to flow along the inner surface of the blade, reducing the pressure difference between the inner and outer sides of the blade, and avoiding the lateral flow of airflow from the high-pressure area to the low-pressure area and the formation of vortices caused by pressure imbalance, thereby reducing the source of vortex noise;

[0089] The presence of the guide channel alters the flow topology inside the blade, dividing the area that would otherwise form large-scale vortices into multiple small-scale, ordered flow regions. These small-scale vortices have lower energy, and the noise frequency components they generate tend to be in the high-frequency range that can be quickly absorbed by the surrounding airflow. Furthermore, the mutual interference between vortices is reduced, decreasing the strong noise pulses generated by vortex interactions and effectively lowering the overall noise level during wind turbine operation.

[0090] Experimental Example 2: Tensile Test (e.g.) Figure 3 (as shown)

[0091] Test equipment: Universal testing machine;

[0092] Test method: The cross-flow fan blades are clamped on a tensile testing machine to check the bonding strength of the fan blade welds and rubber bushings;

[0093] Test requirements: The bonding strength of the fan blade weld and the rubber bushing shall meet the requirements of the drawings or relevant technical documents. If there are no requirements, the fan blade weld shall be able to withstand a tensile force of 4000N axially (9 clamps required).

[0094] The test results are as follows Figure 3 As shown, the tensile test verified that the welded joints of the cross-flow fan blades and the rubber bushing exhibit excellent bonding strength.

[0095] During the tests, the maximum axial tensile forces at the welded joints of the wind turbine blades reached 5490N and 5831N respectively, both far exceeding the technical requirement of 4000N. This indicates that the joint strength has a significant margin and can effectively resist axial tensile forces. Although the deformation varied under different test conditions (7.0896mm and 3.1239mm), no loosening or separation occurred at the welded joints or the rubber bushing joints during the entire test process (85.4 seconds and 37.8 seconds). This fully verifies that the joint strength meets the requirements of the drawings and technical documents, ensuring the structural stability and reliability of the wind turbine blades in actual applications and providing a solid guarantee for the long-term stable operation of the product.

[0096] The above embodiments are only used to illustrate the technical solutions of this utility model, and are not intended to limit it. Although this utility model has been described in detail with reference to the foregoing embodiments, those skilled in the art should understand that modifications can still be made to the technical solutions described in the foregoing embodiments, or equivalent substitutions can be made to some of the technical features. Such modifications or substitutions will not cause the essence of the corresponding technical solutions to deviate from the spirit and scope of the technical solutions of the embodiments of this utility model.

Claims

1. A multi-blade noise-reducing cross-flow fan blade, characterized in that, include: The hub (1) and the central shaft (2) are connected to the hub (1) by a rubber bushing (4); Blades (3) are evenly distributed around the circumference of the hub (1), and the number of blades (3) is 37. The blade (3) has an asymmetric airfoil shape, and the ratio of the radius of curvature R1 of its leading edge to the radius of curvature R2 of its trailing edge is 1:1.2-1.

5. The root of the blade (3) is connected to the hub (1) by welding. The surface of the blade (3) is provided with a microporous noise reduction structure, which is distributed in a gradient along the chord length of the blade.

2. The multi-blade noise-reducing cross-flow fan blade according to claim 1, characterized in that, The blade (3) is made of glass fiber reinforced nylon composite material, and the inner side of the blade is provided with an arc-shaped guide groove (5).

3. The multi-blade noise-reducing cross-flow fan blade according to claim 2, characterized in that, The contact surface between the rubber bushing (4) and the central shaft (2) is provided with a spiral groove.

4. The multi-blade noise-reducing cross-flow fan blade according to claim 3, characterized in that, The asymmetric airfoil has a rounded leading edge.

5. The multi-blade noise-reducing cross-flow fan blade according to claim 4, characterized in that, The gradient density of the microporous noise reduction structure varies with the radial position of the blade.

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