A wet clutch chatter characteristic testing device and testing method
By constructing a wet clutch vibration characteristic testing device, and adopting a fluid isolation distribution design with a dual-channel hollow input shaft and a hydraulic rotary joint, combined with radial oil injection holes and active damping compensation, the problem of inaccurate evaluation of wet clutch vibration characteristics in the prior art is solved, and high-precision vibration signal extraction and evaluation are achieved, supporting the optimization of NVH performance of automatic transmissions.
Patent Information
- Authority / Receiving Office
- CN · China
- Patent Type
- Applications(China)
- Current Assignee / Owner
- HEFEI UNIV OF TECH
- Filing Date
- 2026-04-29
- Publication Date
- 2026-06-23
Smart Images
Figure CN122259218A_ABST
Abstract
Description
Technical Field
[0001] This invention relates to the field of wet clutch technology testing, and more particularly to a wet clutch vibration characteristic testing device and test method. Background Technology
[0002] With the automotive industry's increasing demands for driving comfort and NVH (noise, vibration, and harshness) performance, wet clutches are widely used in automatic transmissions (AT) and wet dual-clutch transmissions (DCT) due to their excellent thermal stability and high torque transmission capability. However, during the slippage stage of engagement, wet clutches are highly susceptible to self-excited vibration, or "chatter," caused by the negative gradient characteristics of the friction coefficient. This low-frequency vibration not only severely degrades the smoothness of vehicle start-up but can also lead to fatigue fracture of transmission components in severe cases, making it a core technical challenge in transmission development.
[0003] Currently, research and testing on clutch vibration characteristics mainly rely on two methods: one is material screening using standard friction testing machines such as SAE No. 2, and the other is conducting real-vehicle road tests. However, standard friction testing machines are mainly used to determine the friction coefficient of friction materials under steady-state conditions and its relationship with speed (μ-v characteristics). Their mechanical structure is designed as a high-rigidity system, which cannot simulate the torsional stiffness and equivalent inertia of a real vehicle's drivetrain. Therefore, they can only predict the vibration trend through the friction slope, but cannot reproduce the actual physical phenomenon of vibration. While real-vehicle tests are realistic, they suffer from high costs, uncontrollable environmental variables, and difficulty in reproducing extreme operating conditions.
[0004] More critically, most existing test benches with dynamic simulation capabilities (such as simulating driveshaft stiffness and vehicle inertia) are developed for dry clutches. Because dry clutches have an open structure and do not require an oil-lubricated environment, their test benches are ill-suited for the complex fluid lubrication, precise oil temperature control, and high-pressure hydraulic actuation conditions required for wet clutches. Currently, the industry lacks a dedicated testing device that integrates wet friction environment control and transmission system dynamic parameter (stiffness, inertia) simulation. This makes it difficult to accurately assess and resolve the vibration problem of wet clutches during the test bench development stage, severely hindering the R&D efficiency of high-performance wet transmission systems.
[0005] Especially during vibration testing and feature identification, the structural vibration of existing testing equipment, the parasitic impedance of the transmission chain, and the inherent mechanical resonance of the test bench often superimpose on the vibration response generated during clutch engagement. Since wet clutch vibration signals often contain high-frequency, low-amplitude dynamic characteristics, these interference components introduced by the test system's background can easily mask the true vibration response. This leads to coupling between the clutch's intrinsic vibration characteristics and the test bench's background response in the acquired signal, thus affecting the accurate identification of key parameters such as vibration frequency, amplitude, and energy distribution. Existing technologies mostly focus on matching friction pair parameters, controlling oil flow, or setting clearances, lacking a testing mechanism for calibrating, separating, and correcting the test bench's background parasitic impedance and mechanical resonance peaks. Therefore, it is difficult to achieve accurate extraction and evaluation of the weak high-frequency vibration signals of wet clutches.
[0006] Furthermore, existing wet clutch technologies and testing equipment still suffer from significant shortcomings in the microscopic control of the lubrication flow field and the dynamic simulation of the load system. On the one hand, existing wet clutch lubrication schemes (such as the wet clutch mechanism disclosed in CN115405639A, which mainly uses fluid guide rings and fluid guide grooves for fluid distribution) are often relatively simple. This conventional fluid distribution method is difficult to overcome the end pressure attenuation caused by centrifugal force and friction resistance when the fluid is transmitted in the hollow shaft, resulting in uneven oil film distribution between friction pairs and failing to meet the precise control requirements for a uniform oil film during the transient process of vibration outbreak. On the other hand, existing clutch test bench calibration and testing methods (such as the calibration method for generating the transfer function of a wet clutch on a test bench disclosed in US20160195451A1) mostly rely on setting initial static physical parameters such as clutch clearance and lubricating oil flow rate. These methods lack an active damping compensation mechanism for high-frequency micro-amplitude vibration during continuous slip friction testing, and therefore cannot realistically and dynamically reproduce the nonlinear damping characteristics of the entire vehicle's transmission system on the test bench.
[0007] Furthermore, on the one hand, there are differences in understanding among those skilled in the art; on the other hand, the applicant studied a large number of documents and patents when making this invention, but due to space limitations, not all details and contents were listed in detail. However, this does not mean that the present invention does not possess the features of these prior art. On the contrary, the present invention already possesses all the features of the prior art, and the applicant reserves the right to add relevant prior art to the background art. Summary of the Invention
[0008] This invention addresses the lack of dedicated testing equipment for wet clutch vibration characteristics in existing technologies, as well as the inability of existing dry test benches to simulate wet fluid environments and standard friction testing machines to simulate the dynamic impedance of real vehicle powertrains. It proposes a wet clutch vibration characteristic testing device and method. This invention aims to achieve a realistic reproduction and objective quantitative evaluation of wet clutch vibration phenomena by constructing a high-fidelity wet thermo-fluid coupling environment and an adjustable dynamic system.
[0009] The wet clutch vibration characteristic testing device disclosed in this invention overcomes the defects of existing technologies through the combination of mechanical spatial layout and closed-loop electronic control logic. Specifically, CN108398262A discloses a wet friction pair torque test bench with basic speed control and data acquisition functions, but it cannot actively dampen the high-frequency micro-vibration of the clutch during continuous sliding friction dynamic process; CN110792648A relates to a hydraulic control system for wet dual clutches, but its fluid distribution circuit lacks a compensation structure for pressure attenuation at the end of the long shaft, resulting in uneven distribution of lubricating oil film at the friction interface; CN202994462U discloses a clutch testing device using an eddy current dynamometer and a magnetic powder loader, but its load application system is in an open-loop or low-frequency response state, making it difficult to cope with the nonlinear high-frequency vibration phenomenon generated by wet clutches under multi-physics coupling.
[0010] To address the technical problems of existing technologies failing to accurately simulate the nonlinear damping characteristics and poor lubrication uniformity of a real vehicle system, this device connects its internal oil guide pipe to a hydraulic rotary joint via a coaxial through-shaft and radial sealing method. This hydraulic rotary joint provides hydraulic actuation and lubrication to a wet clutch assembly with a dual-channel hollow input shaft, enabling fluid isolation and distribution between the hydraulic rotary joint and the internal central cavity and external annular gap of the internal oil guide pipe. Specifically, in conjunction with the attached... Figure 2 The mechanical assembly structure shown has the control oil passage located at the outer annular gap position of the dual-channel hollow input shaft when high-pressure actuation control oil is introduced, so that the outer control oil passage is connected to the sealed pressure chamber behind the clutch piston to provide axial thrust; the lubrication oil passage is located at the central cavity position of the internal oil guide tube when low-pressure cooling lubricating oil is introduced, so that the central lubrication oil passage is connected to the radial oil injection holes arranged in a specific area of the wall of the dual-channel hollow input shaft to achieve lubrication and guidance of low-pressure fluid.
[0011] To address the control requirements of the complex flow field inside a wet clutch assembly, the wall of the dual-channel hollow input shaft extending into the clutch hub is equipped with radial oil injection holes d1, d2, and d3. Under conditions of high-frequency continuous sliding friction and high-speed centrifugal jetting of lubricating oil, the radial oil injection holes d1 and d3 form a specific spatial angle θ along their axial profile, guiding the cooling lubricating oil to the deep sliding friction areas inside the friction plate assembly. This utilizes fluid inertia to reduce the tangential impact when the oil enters the friction plate gap, preventing micro-atomization. Simultaneously, along the axial direction of the dual-channel hollow input shaft, the radial oil injection hole d1, closer to the clutch piston front, has a smaller diameter; the radial oil injection hole d2, in the middle, has a medium diameter; and the radial oil injection hole d3, further from the piston end, has the largest diameter, forming a gradient distribution structure. This structure possesses fluid pressure equalization characteristics with self-compensating axial pressure drop along the fluid transmission direction, overcoming the problem of insufficient end-point oil supply caused by the equal-diameter distribution of oil outlet holes in prior art document CN108398262A.
[0012] In terms of the measurement and control system, the measurement and load simulation unit guides the minute torsional vibration signal transmitted from the end of the solid output shaft to the sensing element of the high-frequency angular acceleration sensor; the dynamic torque measurement component is used to capture the dynamic torque change of the series shaft system in real time; the host computer is used to execute the fast Fourier transform analysis and model comparison of the measurement and control commands. Under the simulated continuous slip friction condition of a real vehicle starting, when the high-frequency angular acceleration sensor senses angular acceleration fluctuations exceeding the noise threshold, it synchronously acquires multi-channel signals with the dynamic torque measurement component; when executing steps S1 to S4 to determine whether vibration has occurred, the dynamic torque measurement component provides negative gradient characteristic data of the friction coefficient changing with the slip velocity at that moment, and performs joint judgment with the host computer; when the preset safety threshold or vibration criterion is triggered, the host computer controls the load motor to output a high-frequency anti-phase compensation torque according to the damping correction model, forming a closed-loop control.
[0013] To ensure the accurate execution of the aforementioned dynamic torque compensation mechanism, the mechanical transmission chain at the rear end of the device possesses high torsional stiffness and zero backlash. When subjected to alternating torque impact, the base plate of the adjustable inertia disk at the output end tightly engages with the shoulder end face of the solid output shaft. The center mating end face of the base plate and the mating shoulder of the solid output shaft achieve high-rigidity meshing and positioning with zero backlash through a Curvic end-face tooth structure. When it is necessary to change the total equivalent inertia J of the system, the counterweight disk aligns and assembles at the circumferential hole on the outer edge of the base plate, achieving a high-rigidity axial fastening connection between the counterweight disk and the base plate. This combination of axial positioning and anti-slip based on end-face teeth provides a stable physical reference for the damping correction model to accurately eliminate the inherent parasitic impedance of the test bench in the frequency domain, supporting the realization of the closed-loop damping compensation function.
[0014] This invention discloses a wet clutch vibration characteristic testing device, comprising a power drive unit, a wet test host unit with a dual-channel hollow input shaft and a wet clutch assembly, a hydraulic power unit with a hydraulic rotary joint, and a measurement and load simulation unit. The dual-channel hollow input shaft has independent control oil passages and lubrication oil passages leading to the wet clutch assembly. The lubrication oil passages have a set of radial oil injection holes at positions corresponding to the wet clutch assembly, with the radial oil injection holes having a preset inclination angle along the rotation direction and exhibiting a gradient distribution in orifice diameter. The measurement and load simulation unit includes a high-frequency angular acceleration sensor, an adjustable inertia disk at the output end, a torsion bar, and a load motor / absorber. The device also includes a host computer configured to receive real-time feedback signals from the high-frequency angular acceleration sensor and output compensation commands based on a built-in damping correction model, driving the load motor / absorber to generate a dynamic torque opposite in phase to the feedback signal, thereby simulating the nonlinear damping characteristics of the target transmission chain.
[0015] According to a preferred embodiment, an internal oil guide tube is coaxially sleeved inside the dual-channel hollow input shaft. The inner cavity of the internal oil guide tube forms a control oil passage, and its outer wall forms a lubrication oil passage with the inner wall of the dual-channel hollow input shaft. The internal oil guide tube is connected to the internal interface of the hydraulic rotary joint through a micro-vibration resistant rotary seal, and the micro-vibration resistant rotary seal has a gap self-compensation mechanism.
[0016] According to a preferred embodiment, the preset tilt angle of the radial injection hole is configured to allow the lubricating oil to cut into the sliding interface in the forward direction; the diameter of the radial injection hole increases sequentially from the side closer to the clutch piston to the side farther away along the axial direction of the dual-channel hollow input shaft, so as to compensate for the pressure drop during lubricating oil transmission by adjusting the orifice pressure drop.
[0017] According to a preferred embodiment, the power drive unit includes a drive motor, which is sequentially connected to an input torque and speed sensor and a simulated inertia flywheel via a coupling; the wet test host unit also includes a sealed wet environment chamber and a solid output shaft, the wet clutch assembly is enclosed inside the wet environment chamber, and the measurement and load simulation unit is connected to the solid output shaft.
[0018] According to a preferred embodiment, the adjustable inertia disk at the output end includes a base disk and a counterweight disk axially pressed onto the base disk. The base disk and the solid output shaft are positioned at the end face by an end face tooth structure. The measurement and load simulation unit also includes a dynamic torque measurement component, and a high-frequency angular acceleration sensor is arranged coaxially adjacent to the dynamic torque measurement component.
[0019] According to a preferred embodiment, the high-frequency angular acceleration sensor adopts a piezoelectric structure, and the host computer also integrates fast Fourier transform signal processing logic to extract the background mechanical noise of the test bench from the signal of the high-frequency angular acceleration sensor by combining a damping correction model.
[0020] The present invention also discloses a method for testing the vibration characteristics of a wet clutch using the aforementioned device, which includes the following steps: S1. Bench background calibration: Perform a frequency sweep test with the wet clutch assembly in the disengaged state, and record the inherent impedance of the bench system to establish a damping correction model. S2. System dynamics matching: Based on the target transmission parameters, physically adjust the mass of the adjustable inertia disk at the output end and / or replace the torsion bar to make the mechanical natural frequency of the test bench match the target vehicle model. S3, Slipping Condition and Dynamic Compensation: Control the wet clutch assembly to generate continuous slipping. During this period, the load motor / absorber outputs dynamic torque in real time according to the damping correction model to perform closed-loop load damping compensation, and simultaneously collects the signal from the high-frequency angular acceleration sensor. S4. Feature Removal and Evaluation: The acquired angular acceleration signal is transformed in the frequency domain. Combined with the damping correction model, the impedance component of the test bench background is removed in the frequency domain to extract pure flutter features and output the evaluation results.
[0021] According to a preferred embodiment, in step S3, the hydraulic power unit supplies high-pressure control oil to the control oil passage via the high-pressure control oil inlet to push the clutch piston to press the friction plate assembly, and simultaneously supplies cooling lubricating oil to the lubrication oil passage via the cooling oil inlet.
[0022] According to a preferred embodiment, in step S3, the drive motor is controlled to rotate at a constant speed, and the load motor / absorber is controlled to apply brakes so that the output speed changes linearly, causing the wet clutch assembly to produce a continuously changing slip speed.
[0023] According to a preferred embodiment, in step S4, the criterion for determining that the clutch is vibrating is: if the power spectral density energy or vibration peak value in a specific vibration-sensitive frequency band exceeds a preset threshold in the frequency domain energy distribution diagram, and the friction coefficient at the same time shows a negative gradient characteristic with the change of slip velocity, then it is determined that there is a risk of vibration.
[0024] Compared with the prior art, the beneficial effects of the present invention are reflected in: 1) Precise control of the microscopic flow field enhances lubrication uniformity: Unlike the simple fluid distribution in existing technologies, this solution designs radial oil injection holes with a preset inclination angle and gradient diameter distribution at the end of the lubrication oil passage of the hollow input shaft. The inclination angle design utilizes fluid inertia to reduce the tangential impact when the lubricating oil enters the friction plate, preventing atomization; the gradient diameter design compensates for axial pressure loss, ensuring that the oil film thickness between each friction pair tends to be uniform. This plays an irreplaceable role in suppressing and objectively evaluating the transient process of chatter outbreaks.
[0025] 2) Active Damping Closed-Loop Compensation for Realistic Reproduction of Nonlinear Dynamic Impedance: Unlike the static parameter matching of existing test benches, this solution achieves active damping compensation through a load motor. Based on real-time feedback from a high-frequency angular acceleration sensor and combined with a damping correction model, the system actively generates a compensation torque with opposite phase, thereby realistically reproducing the nonlinear damping characteristics of the entire vehicle's transmission system on a physical test bench. This completely solves the defects of traditional standard friction testing machine systems, such as fixed impedance and excessive stiffness.
[0026] 3) Modular Design with Adjustable Inertia and Variable Stiffness: By physically replacing torsion bars of different stiffness or adjusting the inertia disk counterweight, and maintaining high-frequency dynamic balance with a high-precision end-face tooth positioning structure, this device possesses the structural basis for adjusting the system's natural frequency within the sensitive resonant frequency range. Researchers can apply these principles of physics... By establishing a vibration model on a test bench that closely resembles the actual vehicle, the technical challenge of predicting vehicle matching risks solely through friction material screening was solved.
[0027] 4) Pure feature stripping based on FFT and model fusion: This invention innovatively introduces a bench background calibration step, uses Fast Fourier Transform (FFT) to extract the energy amplitude of the angular acceleration signal within a specific frequency range, and combines it with a damping correction model to effectively strip away the mechanical resonance peak and friction impedance component of the bench itself, thereby obtaining a true and pure vibration signal of the clutch, which greatly improves the high-frequency testing accuracy of the bench.
[0028] 5) Objective Quantitative Evaluation Mechanism: By combining a high-frequency angular acceleration sensor and a dynamic torque measurement component, the quantitative relationship between the negative gradient characteristics of the clutch friction coefficient and the torsional vibration response of the transmission system was clarified. Existing methods largely rely on the driver's subjective perception. This invention, by monitoring the minute angular velocity fluctuations at the output end in real time, can directly quantify the "vibration intensity" of the vehicle, providing a reliable basis for calibrating automatic transmission anti-vibration control strategies. Attached Figure Description
[0029] Figure 1 This is a schematic diagram of the wet clutch vibration characteristic testing device of the present invention; Figure 2 This is a schematic diagram of the hydraulic actuation and lubrication circuit of the wet clutch assembly in this invention. Figure 3 Microscopic design drawing of radial fuel injection hole; Figure 4 Detailed diagram of the positioning and balancing of the adjustable inertia disk; Figure 5 This is a schematic flowchart of the wet clutch vibration characteristic test method of the present invention; Figure 6Includes: (a) a block diagram of the signal acquisition and input principle of the device of the present invention, (b) a block diagram of the control output execution principle, and (c) a logic diagram of dynamic torque compensation and FFT signal processing; Figure 7 Define the friction coefficient gradient diagram.
[0030] List of reference numerals 1: Drive motor; 2: Input torque and speed sensor; 3: Analog inertia flywheel; 4: Dual-channel hollow input shaft; 4a: Control oil passage; 4b: Lubricating oil passage; 5: Hydraulic rotary joint; 6: Wet clutch assembly; 6a: Return spring; 6b: Clutch piston; 6c: Friction plate assembly; 7: Wet environment chamber; 8: Solid output shaft; 9: Dynamic torque measurement component; 10: High-frequency angular acceleration sensor; 11: Adjustable inertia disk at the output end; 12: Torque bar; 13: Load motor; 14: Hydraulic power unit; 15: High-pressure outlet; 16: Low-pressure inlet; 17: Return oil hose; 18: High-pressure control oil inlet; 19: Cooling oil inlet. Detailed Implementation
[0031] The following is a detailed explanation with reference to the accompanying drawings.
[0032] like Figure 1 As shown, this invention provides a wet clutch vibration characteristic testing device, mainly composed of a power drive unit, a wet testing host unit, a measurement and load simulation unit, and a hydraulic power unit 14. The hydraulic power unit 14 is equipped with a high-pressure outlet 15 and a low-pressure inlet 16. The high-pressure outlet 15 is connected to the control oil passage 4a and / or lubrication oil passage 4b within the dual-channel hollow input shaft 4, used to supply pressurized oil to the system. The low-pressure inlet 16 is connected to the return oil pipeline or oil tank, used to recover low-pressure oil from the system, which is then pressurized by the hydraulic power unit 14 and output from the high-pressure outlet 15, realizing the circulation supply of oil. The entire device is mounted on a cast iron platform with vibration-damping air cushions, i.e., a vibration-damping base, and multiple support bearing seats are provided on the base to stably support the slender shaft structure and isolate environmental vibration from interference with high-frequency vibration signals.
[0033] The power drive unit provides input power to the simulated engine. The power drive unit includes a drive motor 1, which is preferably a low-inertia, high-response AC variable frequency motor. The output shaft of the drive motor 1 is connected to an input torque and speed sensor 2 via a high-rigidity coupling with a flange, for monitoring the real-time operating conditions at the input end. Immediately following the input torque and speed sensor 2 is a simulated inertia flywheel 3, which can be replaced or adjusted by weight according to the engine inertia of the vehicle under test.
[0034] The wet clutch test unit is used to simulate the hydraulic actuation and lubrication environment of a wet clutch. For example... Figure 2As shown, the dual-channel hollow input shaft 4 employs a tube-in-tube structure, containing a coaxially fitted internal oil guide tube. This structure divides the internal space of the shaft into a central, non-communicating control oil passage 4a and an annular lubrication oil passage 4b. A hydraulic rotary joint 5 is installed at the left end of the dual-channel hollow input shaft 4, and has a stator and a rotor; the stator has a high-pressure control oil inlet 18 and a cooling oil inlet 19, which connect to the two oil passages within the shaft, respectively. To prevent oil leakage under high-frequency, low-amplitude torsional vibration conditions, a low-vibration resistant rotary seal is installed between the internal oil guide tube and the internal interface of the hydraulic rotary joint 5. The low-vibration resistant rotary seal uses a PTFE composite lip with a gap self-compensation mechanism, and works with an oil seal spring to provide radial preload, ensuring absolute isolation between the high-pressure control oil passage 4a and the annular lubrication oil passage 4b. A wet clutch assembly 6 is installed at the right end of the dual-channel hollow input shaft 4 and is enclosed within a sealed wet environment chamber 7. The clutch drive hub is fixed to the dual-channel hollow input shaft 4 via a spline. Internally, it integrates a return spring 6a and a clutch piston 6b. The return spring 6a consists of multiple helical springs evenly distributed circumferentially along the drive hub to provide a smooth axial return force. The friction plate assembly 6c is located between the drive hub and the driven hub.
[0035] In terms of actuation and lubrication, high-pressure control oil is transmitted axially through the central control oil passage 4a and enters the pressure chamber behind the clutch piston 6b via an L-shaped radial branch, pushing the clutch piston 6b to overcome the resistance of the return spring 6a and press against the friction plate assembly 6c. Cooling lubricating oil flows to the clutch mounting position through the annular lubrication oil passage 4b. Figure 3 As shown, a radial oil injection hole is radially penetrated through the shaft wall, and the radial oil injection hole has a preset tilt angle θ along the rotation direction. In this embodiment, the preset tilt angle θ is preferably in the range of 30 degrees to 45 degrees. Setting this angle range allows the oil to cut into the high-frequency sliding interface in the forward direction, effectively preventing the lubricating oil from atomizing at the moment of injection, thereby ensuring continuous coverage of the lubricating oil film on the friction plate surface. In addition, the diameter of the radial oil injection hole is gradient distributed along the axial direction from the end near the clutch piston 6b to the end away from it, and the diameter difference between adjacent radial oil injection holes is preferably 0.5 mm, for example, d1 is 2.0 mm, d2 is 2.5 mm, and d3 is 3.0 mm, which satisfies the requirements of the specified parameters. Furthermore, each radial oil injection hole has a chamfered structure at the inlet of the inner wall near the annular lubrication channel 4b to reduce local resistance when the fluid enters the injection hole. This stepped and chamfered design aims to compensate for the end pressure attenuation generated when the lubricating oil is transmitted in the hollow shaft through orifice pressure drop adjustment, ensuring that each layer of friction plates receives a uniform cooling flow. The bottom of the wet environment tank 7 is funnel-shaped and equipped with an oil collection port, which is connected to a large-diameter return oil hose 17, allowing the thrown-out hot oil to smoothly collect and flow back to the oil tank of the hydraulic power unit 14.
[0036] The measurement and load simulation unit is located on the right side of the wet environment chamber 7 in the power output direction to reproduce the dynamic characteristics of the entire vehicle and capture vibration signals. A solid output shaft 8 passes through a sealing mechanism on the right side of the wet environment chamber 7 and connects to the clutch driven hub. This part uses a solid structure to maximize torsional stiffness, ensuring no attenuation occurs when transmitting high-frequency vibration signals. Adjacent to the solid output shaft 8, a dynamic torque measurement assembly 9 and a high-frequency angular acceleration sensor 10 are coaxially mounted via a flange. The high-frequency angular acceleration sensor 10 is a key feature, capable of directly acquiring angular acceleration fluctuations in the shaft system; compared to rotational speed calculus calculations, it is more sensitive to the quantitative characterization of vibration.
[0037] The adjustable inertia disk 11 at the output end and the torsion bar 12 constitute a tunable vibration system. For example... Figure 4 As shown in the detailed and front views, the adjustable inertia disk 11 at the output end includes a base disk and a counterweight disk. The base disk has an internal spline hole at its center, which mates with the external spline at the end of the solid output shaft 8 to transmit torque. The base disk has a thickened outer rim to maximize the base rotational inertia. Multiple mounting holes are evenly distributed circumferentially on its surface for securing the counterweight disk to the base disk with fasteners. Simultaneously, the base disk is fitted onto the external spline at the end of the solid output shaft 8 through the central internal spline hole, with the corresponding end face near the shoulder of the solid output shaft 8 serving as the mounting reference. A Curvic end face tooth or equally spaced trapezoidal tooth structure is provided at this reference end face to achieve circumferential phasing and end face positioning. The counterweight disk is fixedly connected to the base disk by fasteners through the evenly distributed mounting holes circumferentially on its surface. An expansion sleeve is fitted onto the mating area between the base plate and the solid output shaft 8. Upon axial tightening, it generates radial expansion to reduce the mating clearance and suppress assembly eccentricity, ensuring a stable connection between the base plate, counterweight plate, and solid output shaft 8. This achieves zero backlash and self-alignment, guaranteeing the dynamic balance stability of the shaft system under high-frequency reciprocating torque impact. The torsion bar 12 has a stepped shaft structure, with precision alignment sleeves at both ends for limiting its position, simulating the torsional stiffness K of the actual vehicle's drive half-shaft. The adjustable inertia disc 11 at the output end simulates the equivalent mass J of the entire vehicle. By replacing the torsion bar 12 with different diameters or adding / removing the counterweight disc, the natural frequency of the system can be changed. This ensures that the system covers the vehicle vibration sensitive area from 5Hz to 20Hz. The load motor 13 (or absorber) is located at the very end and, in addition to providing basic road driving resistance torque, is connected to a dynamic compensation controller. The dynamic compensation controller receives real-time feedback signals from the high-frequency angular acceleration sensor 10 and actively generates a compensation torque with opposite phase, thereby physically reproducing the nonlinear damping characteristics of the vehicle's drivetrain on a test bench.
[0038] like Figure 6 As shown, the measurement and control system of the present invention is equipped with a host computer. At the signal input end, as... Figure 6As shown in (a), the measurement and control system adopts a parallel signal acquisition architecture. The input terminals of the host computer are independently connected to the input torque and speed sensor 2, the dynamic torque measurement component 9, and the high-frequency angular acceleration sensor 10, respectively. The feedback signals of the above three main sensors are collected in real time and synchronously. At the same time, the oil temperature and oil pressure signals of the hydraulic system are also collected synchronously to ensure the real-time performance and consistency of data acquisition in the time dimension.
[0039] At the control output terminal, such as Figure 6 As shown in (b), the host computer executes commands through three independent control links. Specifically, the output of the host computer is connected to the drive motor controller, the load motor controller, and the hydraulic servo control module, respectively. The control commands generated by the host computer are first sent to each intermediate layer controller. The drive motor controller parses the commands and controls the drive motor 1, and the load motor controller parses the commands and controls the load motor 13, thereby achieving high-precision closed-loop control of speed and torque. Similarly, the host computer sends pressure adjustment commands to the hydraulic servo control module, which then drives the hydraulic power unit 14, thereby achieving precise servo control of the clamping force of the wet clutch assembly.
[0040] In particular, such as Figure 6 As shown in (c), a dynamic torque compensation logic link based on linear feedforward is established within the measurement and control system. This link strictly follows a sequential data processing flow: first, the real-time feedback signal from the high-frequency angular acceleration sensor 10 is acquired; then, this signal is directly input into the built-in damping correction model for processing; next, the compensation torque calculation step is performed based on the output result of the damping correction model; finally, a dedicated load motor control command is generated based on the calculation result. Through this rigorous sequential operation link, the load motor 13 is driven to generate a dynamic torque with a phase opposite to the real-time feedback signal, thereby actively and physically reproducing the nonlinear elastic damping characteristics of the vehicle's transmission chain on the test bench. Simultaneously, the host computer integrates Fast Fourier Transform (FFT) signal processing logic, and the signal processing unit performs FFT processing on the acquired raw high-frequency angular acceleration data. Combined with the aforementioned damping correction model, the background mechanical noise and frictional impedance components of the test bench are extracted from the signal of the high-frequency angular acceleration sensor 10, thereby extracting the pure energy peak amplitude within a specific vibration frequency band, providing accurate data support for subsequent objective quantitative evaluation.
[0041] The test method for wet clutch vibration characteristics using the above-mentioned device specifically includes the following steps: Step S1: Bench Baseline Characteristic Calibration. With the wet clutch assembly 6 in the disengaged state, drive motor 1 is operated and frequency sweep is performed across the entire speed range. The dynamic torque measurement component 9 is used to record the bearing friction torque, air resistance torque, and system parasitic impedance of the bench itself, establishing a damping correction model database to provide benchmark data for subsequent elimination of bench interference.
[0042] Step S2: System Dynamics Matching. Calculate the first-order torsional resonance frequency based on the rotational inertia parameters and transmission parameters of the vehicle under test. Adjust the number of counterweights to change the total mass of the adjustable inertia disc 11 at the output end, and replace the torsion bar 12 with one of the corresponding specifications according to the stiffness of the actual vehicle's transmission system, so that the natural frequency of the bench mechanical system accurately matches the target vehicle model.
[0043] Step S3: Applying slip friction conditions and acquiring data. Start the hydraulic power unit 14 to circulate and heat the ATF oil to a preset temperature, for example, 90°C. Open the cooling and lubrication oil circuit, allowing the cooling oil to circulate back through the cooling oil inlet 19, the annular lubrication oil passage 4b, the friction plate assembly 6c, and the wet environment chamber 7, before returning to the clutch via the return hose 17 to preheat it. Subsequently, control the drive motor 1 to rotate at a constant speed, for example, 1000 rpm, and apply braking through the load motor 13. Open the high-pressure control oil circuit, allowing the hydraulic control oil to be pressurized through the hydraulic rotary joint 5 and enter the high-pressure control oil inlet 18 and the central control oil passage 4a, pushing the clutch piston 6b to overcome the resistance of the return spring 6a and engage the friction plate assembly 6c. Control the load motor 13 to linearly change the output speed, causing the clutch to produce a continuously changing slip speed, for example, scanning from 0 m / s to 2 m / s. During this slip-friction condition, the load motor 13 performs closed-loop torque feedback based on the damping correction model established in step S1, adjusting the load damping at the output shaft end in real time to ensure that the torsional load borne by the friction pair during the test is consistent with the torsional excitation under actual vehicle operating conditions. During this process, the host computer synchronously acquires the torque signal from the dynamic torque measurement component 9, the angular acceleration signal from the high-frequency angular acceleration sensor 10, as well as the input and output speed and oil temperature signals at high speed.
[0044] Step S4: Advanced Feature Extraction and Clean Feature Removal. The signal processing unit in the host computer performs a Fast Fourier Transform on the acquired angular acceleration signal to obtain a frequency domain energy distribution map. Combined with the damping correction model from Step S1, the mechanical resonance peaks and friction impedance components of the test bench itself are effectively removed in the frequency domain, and the negative gradient value of the friction coefficient is calculated. And extract the peak energy amplitude within a specific flutter frequency range.
[0045] Step S5: Flutter evaluation and grade output, and operating condition iteration. The processor is based on the friction coefficient calculated in real time. and the coefficient of friction with sliding velocity The derivative is the friction coefficient gradient. Make a judgment. For example... Figure 7 As shown, based on the mathematical characteristics of the friction coefficient gradient, the system is defined... This represents the stable region corresponding to the positive gradient. The gradient is zero. This represents the flutter region corresponding to a negative gradient. For example... Figure 5 As shown in the judgment logic, if the system detects that the power spectral density energy or angular acceleration vibration peak value in a specific vibration-sensitive frequency range, such as 10Hz to 100Hz, exceeds the preset comfort threshold in the frequency domain energy distribution map. For example And the coefficient of friction at the same moment satisfies That is, fall into Figure 7 Within the defined flutter zone, the friction point is considered to have a flutter risk. Among these, the comfort threshold... The preferred method is to pre-calibrate based on the NVH evaluation standards for the target vehicle model. This involves establishing a mapping relationship between the subjective vibration perception level or the overall vehicle NVH evaluation results and the effective value, peak value, or power spectral density energy of the target frequency band measured under the corresponding operating conditions. This mapping is then converted using a calibration mapping table, a fitting model, or a grading database. Based on the extracted peak energy amplitude and the negative gradient of the friction coefficient, and referring to the NVH evaluation standard database for the target vehicle model, the system outputs a grading result or alarm signal for the clutch vibration tendency. Conversely, if the system determines that no vibration occurs under the current operating condition, it changes the test conditions, such as adjusting parameters like the slip speed range, oil pressure, or temperature, and returns to step S3 to continue testing, thus achieving comprehensive coverage of the multi-condition matrix.
[0046] Step S6: Depressurization and Return. After the test, the hydraulic system executes the depressurization command. Under the elastic force of the return spring 6a, the clutch piston 6b returns to its original position, and the oil in the piston chamber flows back to the hydraulic station oil tank via the central control oil passage 4a, thus completely disengaging the clutch.
[0047] Through the above-described experimental methods and procedures, this invention can realistically and dynamically simulate the thermo-liquid-solid coupling environment and dynamic damping response under actual vehicle operating conditions, providing a high-fidelity, quantifiable, and reliable testing method for optimizing the NVH performance of wet clutches.
[0048] It should be noted that the specific embodiments described above are exemplary. Those skilled in the art can devise various solutions inspired by the disclosure of this invention, and these solutions all fall within the scope of this invention and its protection. Those skilled in the art should understand that this specification and its accompanying drawings are illustrative and not intended to limit the scope of the claims. The scope of protection of this invention is defined by the claims and their equivalents. This specification contains multiple inventive concepts; terms such as "preferredly," "according to a preferred embodiment," or "optionally" indicate that the corresponding paragraph discloses an independent concept. The applicant reserves the right to file divisional applications based on each inventive concept.
Claims
1. A wet clutch vibration characteristic testing device, comprising a power drive unit, a wet test host unit having a dual-channel hollow input shaft (4) and a wet clutch assembly (6), a hydraulic power unit (14) having a hydraulic rotary joint (5), and a measurement and load simulation unit; characterized in that, The dual-channel hollow input shaft (4) is provided with a control oil passage (4a) and a lubrication oil passage (4b) that lead independently to the wet clutch assembly (6). The lubricating oil passage (4b) has a set of radial oil injection holes at the position corresponding to the wet clutch assembly (6). The radial oil injection holes have a preset inclination angle along the rotation direction and are distributed in a gradient of hole diameter. The measurement and load simulation unit includes a high-frequency angular acceleration sensor (10), an adjustable inertia disk (11) at the output end, a torsion bar (12), and a load motor (13). The device also includes a host computer configured to receive the real-time feedback signal from the high-frequency angular acceleration sensor (10) and output compensation commands based on the built-in damping correction model to drive the load motor (13) to generate a dynamic torque opposite to the phase of the feedback signal, so as to simulate the nonlinear damping characteristics of the target transmission chain.
2. The wet clutch vibration characteristic testing device according to claim 1, characterized in that, The internal oil guide pipe is coaxially sleeved inside the dual-channel hollow input shaft (4). The inner cavity of the internal oil guide pipe forms the control oil passage (4a), and its outer wall forms the lubrication oil passage (4b) between the inner wall of the dual-channel hollow input shaft (4). The internal oil guide pipe is connected to the internal interface of the hydraulic rotary joint (5) through a micro-vibration resistant rotary seal, and the micro-vibration resistant rotary seal has a gap self-compensation mechanism.
3. The wet clutch vibration characteristic testing device according to claim 1, characterized in that, The preset angle of the radial oil injection hole is configured to allow the lubricating oil to cut into the sliding interface in the forward direction; the diameter of the radial oil injection hole increases sequentially from the side closer to the clutch piston (6b) to the side farther away along the axial direction of the dual-channel hollow input shaft (4) so as to compensate for the pressure drop during the transmission of lubricating oil by adjusting the orifice pressure drop.
4. The wet clutch vibration characteristic testing device according to claim 1, characterized in that, The power drive unit includes a drive motor (1), which is connected in sequence to an input torque and speed sensor (2) and a simulated inertia flywheel (3) via a coupling; the wet test host unit also includes a sealed wet environment chamber (7) and a solid output shaft (8), the wet clutch assembly (6) is enclosed in the wet environment chamber (7), and the measurement and load simulation unit is connected to the solid output shaft (8).
5. The wet clutch vibration characteristic testing device according to claim 4, characterized in that, The adjustable inertia disk (11) at the output end includes a base disk and a counterweight disk axially pressed onto the base disk. The base disk and the solid output shaft (8) are positioned at the end face through an end face tooth structure. The measurement and load simulation unit also includes a dynamic torque measurement component (9). The high-frequency angular acceleration sensor (10) is arranged coaxially adjacent to the dynamic torque measurement component (9).
6. The wet clutch vibration characteristic testing device according to claim 1, characterized in that, The high-frequency angular acceleration sensor (10) adopts a piezoelectric structure. The host computer also integrates fast Fourier transform signal processing logic, which is used to extract the background mechanical noise of the test bench from the signal of the high-frequency angular acceleration sensor (10) in combination with the damping correction model.
7. A method for testing the vibration characteristics of a wet clutch using the apparatus described in any one of claims 1 to 6, characterized in that, Includes the following steps: S1. Bench background calibration: When the wet clutch assembly (6) is in the disengaged state, a frequency sweep test is performed, and the inherent impedance of the bench system is recorded to establish the damping correction model. S2. System dynamics matching: Based on the target transmission parameters, physically adjust the mass of the adjustable inertia disk (11) at the output end and / or replace the torsion bar (12) so that the mechanical natural frequency of the test bench matches the target vehicle model; S3, Slipping Condition and Dynamic Compensation: Control the wet clutch group (6) to generate continuous slipping. During this period, the load motor (13) outputs dynamic torque in real time according to the damping correction model to perform closed-loop load damping compensation, and simultaneously collects the signal of the high-frequency angular acceleration sensor (10). S4. Feature stripping and evaluation: The acquired angular acceleration signal is transformed in the frequency domain, and the impedance component of the test bench background is removed in the frequency domain by combining the damping correction model, and the pure flutter features are extracted and the evaluation results are output.
8. The test method for the vibration characteristics of a wet clutch according to claim 7, characterized in that, In step S3, the hydraulic power unit (14) supplies high-pressure control oil to the control oil passage (4a) via the high-pressure control oil inlet (18) to push the clutch piston (6b) to press the friction plate assembly (6c), and simultaneously supplies cooling lubricating oil to the lubrication oil passage (4b) via the cooling oil inlet (19).
9. The test method for the vibration characteristics of a wet clutch according to claim 7, characterized in that, In step S3, the drive motor (1) is controlled to rotate at a constant speed, and the load motor (13) is controlled to apply brakes so that the output speed changes linearly, causing the wet clutch assembly (6) to produce a continuously changing slip speed.
10. The test method for the vibration characteristics of a wet clutch according to claim 7, characterized in that, In step S4, the criterion for determining that the clutch is vibrating is: if the power spectral density energy or vibration peak value in a specific vibration-sensitive frequency band exceeds a preset threshold in the frequency domain energy distribution diagram, and the friction coefficient at the same time shows a negative gradient characteristic with the change of slip velocity, then it is determined that there is a risk of vibration.