Blade dynamic loading method and device based on biaxial coupling resonance

Abstract

The invention relates to the technical field of wind power blade fatigue tests, and discloses a blade dynamic loading method based on double-shaft coupling resonance, which comprises the steps of forced loading in the shimmy direction, pulse type follow-up loading in the flapping direction and phase synchronization, and is characterized in that the loading frequency in the shimmy direction is set to be always close to the current resonant frequency in the flapping direction; the amplitude of the exciting force in the shimmy direction is dynamically adjusted according to the vibration response in the flapping direction, in the thrust loading stage, when the displacement in the flapping direction reaches a peak value, a thrust pulse is applied to the flapping direction, the magnitude of the thrust pulse is decreased progressively along with attenuation of the vibration amplitude, energy needed by active excitation is reduced through the modal coupling effect, and the active excitation effect is improved. Pulse thrust is only applied in a displacement peak value interval in the flapping direction, free attenuation is achieved in other time, invalid energy consumption is avoided, the amplitude is adjusted in a self-adaptive mode, so that shimmy excitation force is dynamically optimized along with flapping vibration response, overload or insufficient excitation is avoided, and rigidity degradation in the blade testing process is adapted.

Description

Technical Field

[0001] This invention relates to the field of wind turbine blade fatigue testing technology, specifically to a dynamic loading method and device for blades based on biaxial coupling resonance. Background Technology

[0002] As the core load-bearing component of wind turbine generators, the fatigue performance of wind turbine blades directly determines the operational safety and service life of the entire unit. Therefore, fatigue testing is a crucial step in the blade research and development and mass production stages. Wind turbine blade fatigue testing employs single-axis loading and dual-axis loading. Single-axis loading involves independently loading the blade in both the flapping and yaw directions. However, in actual service, blades simultaneously bear multi-directional composite loads such as wind loads and gravity. The vibrations in the flapping and yaw directions exhibit a coupling effect, and the fatigue damage is the cumulative result of dual-axis coupling. Single-axis loading can only simulate the stress state in a single direction and cannot reproduce the fatigue damage mechanism of dual-axis coupling, leading to a significant deviation between the fatigue life obtained from the test and the actual service life of the blade.

[0003] Biaxial loading involves simultaneously applying dynamic loads to both the flapping and yaw directions of the blade to simulate the coupling effect of multi-directional composite loads during actual blade service. However, during the biaxial loading coupled resonance process, the vibration modes in the flapping and yaw directions can cause cross-interference, requiring additional energy to suppress this coupling interference. This not only significantly increases the testing cost but also reduces the stability of the system operation. To address these issues, we propose a dynamic loading method and device for blades based on biaxial coupled resonance. Summary of the Invention

[0004] The present invention aims to provide a dynamic loading method and apparatus for blades based on biaxial coupling resonance, so as to solve the problem that biaxial loading will cause cross interference, and additional energy needs to be consumed to suppress this coupling interference.

[0005] To achieve the above objectives, the present invention adopts the following technical solution: a blade dynamic loading method based on biaxial coupling resonance, used to control a blade dynamic loading device based on biaxial coupling resonance, characterized in that it includes forced loading in the flapping direction, pulse-type follow-up loading in the flapping direction, and phase synchronization. The forced loading in the flapping direction includes: Resonance frequency tracking, the applied frequency (f) in the oscillation direction. edge ) Always close to the current resonant frequency (f) in the direction of the swing flap That is, it satisfies the following relation: ; Amplitude adaptive adjustment, the amplitude of the excitation force (F) in the oscillation direction edge The dynamic adjustment is based on the vibration response in the direction of the swing, satisfying the following relationship: A flapThe real-time amplitude is the waving direction, k is the gain coefficient, and F is the amplitude. base Basic excitation force; The waving direction pulse-type follow-up loading includes: During the thrust loading phase, when the displacement in the waving direction reaches its peak, a thrust pulse is applied in the waving direction. The magnitude of the thrust pulse decreases as the vibration amplitude decays, satisfying the following relationship: e is the natural constant, and λ is the attenuation coefficient; During the free decay phase, after the thrust pulse is applied, the blade dynamic loading device enters a low-pressure standby mode, allowing the blade to vibrate freely. Phase synchronization includes: Phase difference locking ensures that the oscillation excitation and the flapping vibration have a 90° phase difference, thus maximizing energy input efficiency. Real-time monitoring and acquisition of vibration signals in the swinging and waving directions, and calculation of the current actual phase difference; Dynamic correction: fine-tuning the oscillation frequency or the timing of the flapping thrust based on the phase difference until the actual phase difference returns to the target value.

[0006] The beneficial effects of this scheme are: the loading frequency in the oscillation direction is based on... Dynamically tracking the wave resonant frequency, utilizing modal coupling effects to reduce the energy required for active excitation, applying pulse thrust only during the peak displacement range in the wave direction, allowing free decay at other times, avoiding unnecessary energy consumption, and adaptively adjusting the amplitude. The relationship allows for dynamic optimization of the wagging excitation force in response to the flapping vibration, avoiding overload or insufficient excitation, adapting to stiffness degradation during blade testing, and ensuring stable phase relationship during biaxial loading to prevent system instability. It effectively suppresses cross-interference between the flapping and wagging directions. Phase synchronization and adaptive amplitude adjustment form a dual dynamic control, effectively avoiding risks such as overload and frequency deviation during loading, ensuring stable system operation. The thrust pulse in the flapping direction is... It exhibits exponential decay characteristics, closely matching the randomness and decay law of real wind loads, further improving the realism of load simulation and providing more reliable data support for blade fatigue performance evaluation.

[0007] Preferably, as an improvement, it also includes decoupling control, which includes cross-interference suppression. It employs a feedforward compensation algorithm to adjust the output force in the swing direction in advance to offset the load in the swing direction.

[0008] The beneficial effects are as follows: By using the feedforward compensation algorithm, based on the pre-modeling relationship between the oscillation load and the waving interference, the non-target interference force of the oscillation direction on the waving direction can be predicted in advance. The interference force and the compensation force can be completely canceled by the output of the waving cylinder, so that only the target pulse thrust is retained in the waving direction. This avoids the distortion of the waving load caused by the oscillation interference and ensures that the dual-axis loading is independent and controllable.

[0009] Preferably, as an improvement, a security protection mechanism is also included, which includes the following steps: Step 1: When the load in any direction exceeds the set threshold, the blade dynamic loading device releases pressure. Step 2: The blade dynamic loading device returns to zero position and stops excitation; Step 3: While depressurizing and stopping the vibration, lock the loading ends in the swinging and waving directions.

[0010] The beneficial effects are as follows: when the oscillation excitation force or flapping pulse thrust exceeds the set threshold, the blade dynamic loading device depressurizes, which can cut off the power source before the load causes micro-cracks in the blade, thus avoiding the continuous increase of load leading to local blade fracture. The blade dynamic loading device returns to the zero position, which on the one hand avoids the superposition of blade inertial vibration caused by untimely vibration cessation, and on the other hand restores the loading device to the initial state, preventing secondary impact caused by mechanism misalignment during subsequent restart. While depressurizing and stopping vibration, the loading ends in the oscillation direction and flapping direction are locked, and the residual vibration generated by blade rebound is eliminated by the rigid constraint of the mechanical brake.

[0011] Preferably, as an improvement, a blade dynamic loading device based on biaxial coupling resonance includes a loading base, a oscillating loading component, a waving loading component, and a blade clamp. The blade clamp is disposed on the outer wall of the blade, the oscillating loading component is disposed on the side wall of the loading base, and the waving loading component is disposed on the upper end of the loading base, with the oscillating loading component and the waving loading component being perpendicular. The oscillating loading component includes an oscillating base, a first driving member, and a first loading arm. The oscillating base is slidably disposed on the side wall of the loading base, and the first loading arm is rotatably disposed at one end of the oscillating base. The first driving member is disposed inside the oscillating base, and the output end of the first driving member is rotatably connected to the first loading arm. The waving loading component includes a waving base, a second driving member, and a second loading arm. The waving base is slidably disposed on the upper end of the loading base, and the second loading arm is rotatably disposed at one end of the waving base. The second driving member is disposed inside the waving base, and the output end of the second driving member is rotatably connected to the second loading arm. Connectors are provided between the first loading arm, the second loading arm, and the blade clamp. A hydraulic system is provided at the upper end of the loading base.

[0012] The beneficial effects are as follows: the oscillation loading component and the swing loading component are arranged perpendicularly, and can drive the first loading arm and the second loading arm to move respectively. This allows the device to perform loading in the swing or oscillation direction alone, as well as to achieve dual-axis coupled loading, meeting the simulation requirements of complex working conditions. Moreover, the first driving component and the second driving component of this solution are rotatably connected to the first loading arm and the second loading arm respectively. By controlling the extension and retraction of the first driving component and the second driving component, the loading force, displacement, frequency and other parameters of the device can be precisely adjusted, so that the blade can be loaded according to the expected swing curve, thereby improving the loading accuracy of the device.

[0013] Preferably, as an improvement, the loading base includes a pump station base frame, a waving base frame, a waving guide rail, a swinging upright frame, and a swinging guide rail. The pump station base frame and the waving base frame are fixedly connected. The hydraulic system is located at the upper end of the pump station base. The waving guide rail is located at the upper end of the waving base frame. The swinging upright frame is located on one side of the upper end of the waving base frame. The swinging guide rail is located at the end of the swinging upright frame away from the waving guide rail, and the swinging guide rail is perpendicular to the waving guide rail. The lower end of the swinging base is provided with a first slider that matches the swinging guide rail. The first slider is slidably installed on the upper end of the swinging guide rail. The lower end of the waving base is provided with a second slider that matches the waving guide rail. The second slider is slidably installed on the upper end of the waving guide rail.

[0014] The beneficial effects are as follows: the oscillating support and the flapping base together form a rigid frame with vertical support and horizontal load bearing, which can directly bear the high load impact during the test of a 150-meter-class large wind turbine blade, avoiding the deviation in loading accuracy caused by base deformation. The flapping base can be adjusted along the width direction of the blade through the cooperation of the second slider and the flapping guide rail, ensuring that the flapping loading force can be applied to the specified loading section of the blade. The oscillating base can be adjusted along the height direction of the blade through the cooperation of the first slider and the oscillating guide rail, adapting to the oscillating loading requirements of blades of different lengths.

[0015] Preferably, as an improvement, the oscillation loading assembly further includes two sets of first adjusting members symmetrically arranged on the side wall of the oscillation stand. The first adjusting member includes a fixed seat, a lead screw, and a lifting turbine. The fixed seat is arranged on the side wall of the oscillation stand, the lead screw is arranged between the fixed seats, the lifting turbine is arranged on the outer wall of the lead screw, and a first motor is arranged on the outer wall of the lifting turbine. The output end of the first motor is connected to the input end of the lifting turbine, and the lifting turbine is fixedly connected to the oscillation base.

[0016] Preferably, as an improvement, the swing loading component further includes a second adjusting member, which includes a chain, a gear, and a second motor. The chain is located at the upper end of the swing base, the second motor is located inside the swing base, and the gear is located at the output end of the second motor, and the gear is matched with the chain.

[0017] Preferably, as an improvement, the first driving component includes a first control valve group and a first servo cylinder. The first control valve group is located at the end of the oscillating base away from the oscillating support, and the first servo cylinder is located inside the oscillating base. The first control valve group is connected to the hydraulic system and the first servo cylinder respectively.

[0018] Preferably, as an improvement, the second driving component includes a second control valve group and a second servo cylinder. The second control valve group is located at the upper end of the swing base, and the second servo cylinder is located inside the swing base. The second control valve group is connected to the hydraulic system and the second servo cylinder respectively.

[0019] The beneficial effects are as follows: by adjusting the hydraulic oil flow of the hydraulic system to the first servo cylinder and the second servo cylinder through the first control valve group and the second control valve group, the extension and retraction speed of the output end of the first servo cylinder and the second servo cylinder are controlled respectively, thereby accurately matching the frequency parameters of oscillation loading and flapping loading, realizing the bending motion trajectory of the blade with dual-axis coupling, greatly improving the dynamic response speed and accuracy of dual-axis loading, and adapting to the fatigue test requirements of large blades with high load and high frequency.

[0020] Preferably, as an improvement, the connector includes a loading link and two mounting seats, one of which is located on the outer wall of the blade clamp and the other is located on the outer wall of the first loading arm or the second loading arm. Both ends of the loading link are provided with universal joints, and the universal joints at both ends of the loading link are rotatably connected to the corresponding mounting seats.

[0021] The beneficial effects are as follows: By using a loading linkage for rigid connection, the mechanical force output by the first and second loading arms can be directly and losslessly transmitted to the blade fixture, reducing test data deviation. At the same time, the loading linkage can also bear the reaction force during high-load tests of large blades, avoiding bending or breakage and ensuring the stability of long-term tests. Furthermore, when the blade undergoes bending deformation under biaxial loading, the universal joints at both ends of the loading linkage flexibly adapt to the angle changes between the first and second loading arms and the blade fixture, avoiding the forced force problem caused by rigid connection, reducing stress concentration in the loading linkage, the first loading arm, the second loading arm, and the blade fixture, and extending the service life of each component. Moreover, by adjusting the distance between the two loading linkages and the hinge points of the first and second loading arms, i.e., the distance between the loading linkages and the loading points of the first and second servo cylinders, the output displacement of the first and second servo cylinders can be amplified to the actual requirements using the lever principle, thereby meeting the fatigue test requirements for large blades. Attached Figure Description

[0022] Figure 1 This is a three-dimensional structural schematic diagram of the biaxial loading device according to an embodiment of the present invention; Figure 2This is a three-dimensional structural diagram of the loading base according to an embodiment of the present invention; Figure 3 This is a schematic diagram of the waving loading component according to an embodiment of the present invention; Figure 4 This is a cross-sectional view of the waving loading component according to an embodiment of the present invention; Figure 5 This is a three-dimensional structural schematic diagram of the oscillation loading component according to an embodiment of the present invention; Figure 6 for Figure 1 A magnified schematic diagram of the structure at point A in the middle. Detailed Implementation

[0023] The following detailed description illustrates the specific implementation method: The reference numerals in the accompanying drawings include: 1. Loading base; 2. Blade; 3. Blade clamp; 4. Pump station base frame; 5. Swinging base frame; 6. Swinging brace; 7. Swinging brace; 8. Hydraulic system; 9. Swinging base; 10. Second loading arm; 11. Second control valve group; 12. Second servo cylinder; 13. Suspension spring; 14. Swinging base; 15. First loading arm; 16. First control valve group; 17. First servo cylinder; 18. Fixed seat; 19. Lead screw; 20. Lifting turbine; 21. Second motor; 22. Loading connecting rod; 23. Mounting seat; 24. Universal joint; 25. Swinging loading assembly; 26. Swinging loading assembly; 27.

[0024] Example The basic implementation examples are as follows: Figures 1-6 As shown, Figure 1 The illustrated blade dynamic loading device based on biaxial coupled resonance includes a loading base 1, a tumbling loading assembly 26, a flapping loading assembly 27, and a blade clamp 3. The blade clamp 3 is fixedly installed on the outer wall of the blade 2. The outer wall of the blade 2 is equipped with strain gauges and a first acceleration sensor for detecting changes in the natural frequency of the blade 2. Figure 2 The loading base 1 shown includes a pump station base frame 4, a waving base frame 5, a waving guide rail, a swing vibration upright 6, and a swing vibration guide rail. The pump station base frame 4 is fixedly installed at the front end of the waving base frame 5. The waving guide rail is symmetrically fixedly installed at the upper end of the waving base frame 5. The swing vibration upright 6 is fixedly installed on the upper left side of the waving base frame 5. A swing vibration diagonal brace 7 is fixedly installed between the front end of the swing vibration upright 6 and the upper end of the pump station base frame 4. A waving diagonal brace 8 is fixedly installed between the right end of the swing vibration upright 6 and the upper end of the waving base frame 5. The swing vibration guide rail is symmetrically fixedly installed at the left end of the swing vibration upright 6, and the swing vibration guide rail is perpendicular to the waving guide rail.

[0025] like Figure 1The pump station base frame 4 shown is fixedly mounted with a hydraulic system 9. The hydraulic system 9 includes a multi-pump pressure supply unit and two oil distributors. The two oil distributors are respectively connected to the first control valve group 17 and the second control valve group 12 through pipelines. Each oil distributor is equipped with a proportional relief valve, such as... Figure 1 The oscillation loading component 26 shown is slidably mounted on the left end of the oscillation support 6, and the swing loading component 27 is slidably mounted on the upper end of the swing support, with the oscillation loading component 26 and the swing loading component 27 perpendicular to each other. Figure 3 The wave loading assembly 27 shown includes a wave base 10, a second drive component, and a second loading arm 11. A second slider matching the wave guide rail is fixedly mounted on the lower end of the wave base 10. The second slider is slidably mounted on the upper end of the wave guide rail. The left end of the second loading arm 11 is rotatably mounted on the right end of the wave base 10 via a bearing. The second drive component includes a second control valve group 12 and a second servo cylinder 13. The second servo cylinder 13 is configured as a rapid pressure relief cylinder and has a built-in displacement sensor, such as... Figure 4 The second control valve assembly 12 is fixedly installed on the upper end of the waving base 10. The right end of the second servo cylinder 13 is rotatably connected to the left wall of the waving base 10, and the output end of the second drive component is rotatably connected to the left end of the second loading arm 11. A suspension spring 14 is rotatably installed between the upper wall of the waving base 10 and the upper end of the second servo cylinder 13. The suspension spring 14 is used to balance the weight of the second servo cylinder 13. Figure 5 The oscillation loading assembly 26 shown includes an oscillation base 15, a first driving component, and a first loading arm 16. A first slider matching the oscillation guide rail is fixedly mounted at the rear end of the oscillation base 15. The first slider is slidably mounted on the upper end of the oscillation guide rail. The first loading arm 16 is rotatably mounted on the upper end of the oscillation base 15 via bearings. The first driving component includes a first control valve group 17 and a first servo cylinder 18. The first servo cylinder 18 is a high-frequency response servo cylinder, which, in conjunction with the first control valve group 17, achieves a forced vibration amplitude of ±0.8m@2Hz. The first control valve assembly 17 is fixedly installed at the front end of the oscillating base 15. The lower end of the first servo cylinder 18 rotates on the lower wall of the oscillating base 15, and the output end of the first servo cylinder 18 is rotatably connected to the lower end of the first loading arm 16. The first control valve assembly 17 is connected to the hydraulic system 9 and the first servo cylinder 18 respectively through pipelines. The second control valve assembly 12 is connected to the hydraulic system 9 and the second servo cylinder 13 respectively through pipelines. The outer walls of the output ends of the first servo cylinder 18 and the second servo cylinder 13 are both fixedly installed with second acceleration sensors.

[0026] like Figure 5The swing loading assembly 26 shown also includes two sets of first adjusting components symmetrically arranged at the front end of the swing support 6. Each first adjusting component includes a fixed base 19, a lead screw 20, and a lifting turbine 21. The fixed base 19 is fixedly installed at the front end of the swing support 6, the lead screw 20 is fixedly installed between the fixed bases 19, and the lifting turbine 21 is threadedly connected to the outer wall of the lead screw 20. A first motor is fixedly installed on the outer wall of the lifting turbine 21, and the output end of the first motor is connected to the input end of the lifting turbine 21. The upper end of the lifting turbine 21 is fixedly connected to the outer wall of the swing base 15 by bolts. The swing loading assembly 27 also includes a second adjusting component, which includes a chain, gears, and a second motor 22. The chain is symmetrically fixedly installed on the upper end of the swing base 5. Figure 4 The second motor 22 shown is fixedly installed at the lower end of the waving base 10. The gear is fixedly installed at the output end of the second motor 22 and is matched with the chain. The second motor 22 drives the gear to rotate the rod, thereby driving the waving base 10 to move along the chain direction.

[0027] like Figure 1 The first loading arm 16, the second loading arm 11, and the blade clamp 3 are all provided with connecting parts. The connecting parts include a loading connecting rod 23 and two mounting seats 24. In this embodiment, taking the second loading arm 11 as an example, one mounting seat 24 is fixedly installed on the upper end of the second loading arm 11 by bolts, and the other mounting seat 24 is fixedly installed on the lower end of the blade clamp 3 by bolts. Figure 6 Universal joints 25 are fixedly installed at both ends of the loading link 23 shown, and the universal joints 25 at both ends of the loading link 23 are rotatably connected to the corresponding mounting bases 24.

[0028] A blade dynamic loading method based on biaxial coupling resonance includes forced loading in the flapping direction, pulse-type follow-up loading in the flapping direction, phase synchronization, decoupling control, and a safety protection mechanism, which is used to control a blade dynamic loading device based on biaxial coupling resonance.

[0029] Forced loading in the oscillation direction includes: Resonance frequency tracking, the applied frequency (f) in the oscillation direction. edge ) Always close to the current resonant frequency (f) in the direction of the swing flap That is, it satisfies the following relation: ; The natural frequency change of blade 2 is detected by strain gauges and a first accelerometer to ensure that the optimal excitation frequency is maintained even when the stiffness of blade 2 degrades (such as due to fatigue damage accumulation).

[0030] Amplitude adaptive adjustment, the amplitude of the excitation force (F) in the oscillation direction edge The vibration response is dynamically adjusted according to the direction of the swing to avoid overload or insufficient excitation, thus satisfying the following relationship: ; Among them, A flap F represents the real-time amplitude in the flapping direction, and k is the gain coefficient. The value of k is determined jointly based on the dynamic characteristics of blade 2 and the control objective. The dynamic characteristics of blade 2 are related to the modal coupling strength, the damping characteristics in the flapping direction, and the stiffness level of blade 2. base The basic excitation force.

[0031] The waving direction pulse-type follow-up loading includes: During the thrust loading phase, when the displacement sensor detects that the displacement in the swing direction has reached its peak, the hydraulic system 9 applies a short-duration thrust pulse, the duration of which is 0.25-0.30 times the current vibration cycle, in order to maintain the resonance state.

[0032] The thrust pulse magnitude decreases as the vibration amplitude decays, simulating the randomness of real wind loads, i.e., satisfying the following relationship: ; Where t is the time when the thrust pulse begins to be applied, and F flap(t) Let F be the pulse thrust applied by the hydraulic system 9 to the blade 2 in the flapping direction at a certain moment, where e is the natural constant, λ is the attenuation coefficient, and is set according to the material damping characteristics. max Apply the initial maximum output force to the thrust.

[0033] During the free decay phase, after the thrust pulse ends, the hydraulic system 9 enters a low-pressure standby mode, allowing the blades 2 to vibrate freely and avoiding additional energy consumption.

[0034] Phase synchronization includes: Phase difference locking ensures that the pendulum excitation and the swinging vibration form a 90° phase difference, meaning that the pendulum force is applied when the swinging speed is at its maximum, in order to maintain maximum energy input efficiency. Real-time monitoring involves collecting vibration signals in the swinging and waving directions using strain gauges and a second accelerometer, and then calculating the current actual phase difference using FFT analysis. Dynamic correction: fine-tuning the oscillation frequency or the timing of the flapping thrust based on the phase difference until the actual phase difference returns to the target value.

[0035] Decoupling control includes: Cross-interference suppression: Due to the coupling of the two-axis motion, a feedforward compensation algorithm is adopted. Based on the load in the swing direction, its influence on the waving direction is predicted, and the output force of the second servo cylinder 13 is adjusted in advance to offset it. The compensation amount can be obtained through experimental calibration or finite element model calculation.

[0036] The security protection mechanism includes the following steps: Step 1: When the load in any direction exceeds the set threshold, each proportional relief valve will release pressure and drop to 5MPa within 20ms, quickly cutting off the power supply to the hydraulic system 9 and preventing the load from continuing to increase. Step 2: The first servo valve group and the second servo valve group return to the zero position and stop the excitation. The first servo cylinder 18 and the second servo cylinder 13 stop the loading movement. Step 3: While depressurizing and stopping vibration, the mechanical brake is activated to lock the output ends of the first servo cylinder 18 and the second servo cylinder 13, eliminating the inertia of the output ends of the first servo cylinder 18 and the second servo cylinder 13 and the residual vibration caused by the rebound of the blade 2, thus avoiding secondary damage. The safety protection mechanism also includes: if the deviation between the swing loading frequency and the swing resonance frequency exceeds ±10%, an alarm will be triggered and the excitation parameters will be automatically adjusted.

[0037] The specific implementation process is as follows: When conducting fatigue testing on blade 2, the blade clamp 3 is first fixedly installed at the test position of blade 2. Then, the device is moved to the test side of blade 2. The positions of the oscillation loading assembly 26 and the flapping loading assembly 27 are adjusted by controlling the first and second adjusting components, respectively. After the oscillation loading assembly 26 and the flapping loading assembly 27 are moved to the designated positions, the operator uses bolts to fix the oscillation base 15 and the oscillation support 6, and the flapping base 10 and the flapping frame 5, respectively. Then, the first loading arm 16 and the second loading arm 11 are connected to the blade clamp 3 using connectors. Then, by activating the hydraulic system 9, hydraulic oil is supplied to the first servo cylinder 18 and the second servo cylinder 13 respectively. Thus, the first servo cylinder 18 and the second servo cylinder 13 drive the first loading arm 16 and the second loading arm 11 to swing around the swinging base 15 and the waving base 10 respectively, thereby causing the blade 2 to swing in the horizontal and vertical directions. By adjusting the loading force, displacement, frequency and other parameters of the first servo cylinder 18 and the second servo cylinder 13, high load, high precision, independent or coupled loading of the swinging direction and waving direction of the blade 2 can be achieved, so that the blade 2 can achieve the target swing curve.

[0038] In this scheme, the loading frequency in the oscillation direction is based on... Dynamically tracking the wave resonant frequency, utilizing modal coupling effects to reduce the energy required for active excitation, applying pulse thrust only during the peak displacement range in the wave direction, allowing free decay at other times, avoiding unnecessary energy consumption, and adaptively adjusting the amplitude. The relationship allows for dynamic optimization of the oscillation excitation force in response to the flapping vibration, avoiding overload or insufficient excitation, adapting to the stiffness degradation during blade 2 testing, and ensuring stable phase relationship during biaxial loading to prevent system instability. It effectively suppresses cross-interference between flapping and oscillation directions. Phase synchronization and amplitude adaptive adjustment form a dual dynamic control, effectively avoiding risks such as overload and frequency deviation during loading, ensuring stable system operation. The thrust pulse in the flapping direction is... It exhibits exponential decay characteristics, which closely match the randomness and decay law of real wind loads, further improving the realism of load simulation and providing more reliable data support for the fatigue performance evaluation of blade 2.

[0039] The above descriptions are merely embodiments of the present invention, and common knowledge such as specific technical solutions and / or characteristics are not described in detail here. It should be noted that those skilled in the art can make various modifications and improvements without departing from the technical solutions of the present invention, and these should also be considered within the scope of protection of the present invention. These modifications and improvements will not affect the effectiveness of the implementation of the present invention or the practicality of the patent. The scope of protection claimed in this application should be determined by the content of its claims, and the specific embodiments described in the specification can be used to interpret the content of the claims.

Claims

1. A method for controlling a dual-axis coupled resonance based blade dynamic loading apparatus, the method comprising: determining a desired blade dynamic loading; determining a desired blade dynamic loading frequency; determining a desired blade dynamic loading phase; and controlling the dual-axis coupled resonance based blade dynamic loading apparatus to achieve the desired blade dynamic loading, the desired blade dynamic loading frequency, and the desired blade dynamic loading phase. The method comprises forced loading in the edgewise direction, impulse servo loading in the flapwise direction, and phase synchronization, the forced loading in the edgewise direction comprising: The resonant frequency tracking, the loading frequency (f edge ) in the pitch direction is always close to the current resonant frequency (f flap ) in the flap direction, i.e. the relationship is satisfied: ; The amplitude is self-adaptively adjusted, and the amplitude of the exciting force in the swing direction (F edge ) is dynamically adjusted according to the vibration response in the flap direction, that is, the following relationship is met: , A flap is the real-time amplitude in the flap direction, k is a gain coefficient, and F base is a basic exciting force; The impulse servo loading in the flapwise direction comprises: In the thrust loading stage, a thrust pulse is applied to the flapwise direction when the displacement reaches the peak value, and the thrust pulse size decreases with the vibration amplitude decay, that is, it satisfies the relationship: , e is the natural constant, and λ is the decay coefficient. A free decay stage, after the end of the thrust pulse, the blade dynamic loading device enters a low-pressure standby mode, allowing the blade to vibrate freely; The phase synchronization comprises: Phase difference locking, the edgewise excitation and the flapwise vibration form a 90° phase difference to maintain maximum energy input efficiency; Real-time monitoring, vibration signals in the edgewise and flapwise directions are collected to calculate the current actual phase difference; Dynamic correction, the edgewise frequency or the flapwise thrust triggering time is adjusted according to the phase difference until the actual phase difference returns to the target value.

2. The method of claim 1, wherein the method is based on a dual-axis coupling resonance. The method further comprises decoupling control, the decoupling control comprising cross-interference suppression, a feedforward compensation algorithm is used to adjust the output force in the flapwise direction in advance to offset the load in the edgewise direction.

3. The method of claim 2, wherein the method is based on a dual-axis coupling resonance. The method further comprises a safety protection mechanism, the safety protection mechanism comprising the following steps: Step 1, when the load in any direction exceeds the set threshold, the blade dynamic loading device is depressurized; Step 2, the blade dynamic loading device returns to zero and stops exciting; Step 3, while depressurizing and stopping, the loading ends in the edgewise and flapwise directions are locked.

4. A dynamic loading device for a blade based on two-axis coupling resonance, characterized in that: The device comprises a loading base, an edgewise loading assembly, a flapwise loading assembly, and a blade clamp, the blade clamp is arranged on the outer wall of the blade, the edgewise loading assembly is arranged on the side wall of the loading base, the flapwise loading assembly is arranged on the upper end of the loading base, and the edgewise loading assembly is perpendicular to the flapwise loading assembly, the edgewise loading assembly comprises an edgewise base, a first driving member, and a first loading arm, the edgewise base is slidingly arranged on the side wall of the loading base, the first loading arm is rotatably arranged at one end of the edgewise base, the first driving member is arranged in the edgewise base, and the output end of the first driving member is rotatably connected with the first loading arm, the flapwise loading assembly comprises a flapwise base, a second driving member, and a second loading arm, the flapwise base is slidingly arranged on the upper end of the loading base, the second loading arm is rotatably arranged at one end of the flapwise base, the second driving member is arranged in the flapwise base, and the output end of the second driving member is rotatably connected with the second loading arm, a connecting member is arranged between the first loading arm, the second loading arm, and the blade clamp, and a hydraulic system is arranged on the upper end of the loading base.

5. The dynamic loading device of claim 4, wherein: The loading base comprises a pump station base frame, a flapwise base frame, a flapwise guide rail, an edgewise stand, and an edgewise guide rail, the pump station base frame is fixedly connected with the flapwise base frame, the hydraulic system is arranged on the upper end of the pump station base, the flapwise guide rail is arranged on the upper end of the flapwise base frame, the edgewise stand is arranged on one side of the upper end of the flapwise base frame, the edgewise guide rail is arranged at the end of the edgewise stand away from the flapwise guide rail, and the edgewise guide rail is perpendicular to the flapwise guide rail, the lower end of the edgewise base is provided with a first sliding block matched with the edgewise guide rail, the first sliding block is slidingly installed on the upper end of the edgewise guide rail, and the lower end of the flapwise base is provided with a second sliding block matched with the flapwise guide rail, the second sliding block is slidingly installed on the upper end of the flapwise guide rail.

6. A dynamic loading device for a blade based on two-axis coupling resonance according to claim 5, characterized in that: The swing loading assembly further comprises two sets of first adjusting members symmetrically arranged on the side walls of the swing stand, each of the first adjusting members comprising a fixing seat, a screw rod and a lifting turbine, the fixing seat being arranged on the side wall of the swing stand, the screw rod being arranged between the fixing seats, the lifting turbine being arranged on the outer wall of the screw rod, the outer wall of the lifting turbine being provided with a first motor, the output end of the first motor being connected with the input end of the lifting turbine, the lifting turbine being fixedly connected with the swing base.

7. A dynamic loading device for a blade based on two-axis coupling resonance according to claim 6, characterized in that: The flap loading assembly further comprises a second adjusting member, the second adjusting member comprising a chain, a gear and a second motor, the chain being arranged on the upper end of the flap base, the second motor being arranged in the flap base, the gear being arranged on the output end of the second motor and matched with the chain.

8. The dynamic loading device of claim 7, wherein: The first driving member comprises a first control valve group and a first servo oil cylinder, the first control valve group being arranged on the end of the swing base away from the swing stand, the first servo oil cylinder being arranged in the swing base, and the first control valve group being connected with the hydraulic system and the first servo oil cylinder respectively.

9. The dynamic loading device of claim 8, wherein: The second driving member comprises a second control valve group and a second servo oil cylinder, the second control valve group being arranged on the upper end of the flap base, the second servo oil cylinder being arranged in the flap base, and the second control valve group being connected with the hydraulic system and the second servo oil cylinder respectively.

10. The bladed dynamic loading device based on two-axis coupling resonance according to claim 9, characterized in that: The connecting member comprises a loading connecting rod and two mounting seats, one of the mounting seats being arranged on the outer wall of the blade clamp, the other mounting seat being arranged on the outer wall of the first loading arm or the second loading arm, the two ends of the loading connecting rod being provided with universal joints, and the universal joints at the two ends of the loading connecting rod being rotatably connected with the corresponding mounting seats.

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