Ackerman self-adaptive correction system and method based on transverse leaf spring end displacement sensing

Abstract

The present application relates to the technical field of corner module, and in particular to an Ackermann self-adaptive correction system and method based on end displacement sensing of a transverse leaf spring, comprising a corner module suspension mechanism, the corner module suspension mechanism comprising a lower control arm, a subframe and a transverse leaf spring as an elastic element, and the free end of the transverse leaf spring being connected to the lower control arm through a non-rigid constraint lapping mode. The present application utilizes the mechanical accompanying deformation feature of the longitudinal relative sliding displacement generated by the nonlinear elongation of the free end chord length when the transverse leaf spring is loaded and bent, and through the multi-element fusion of the vertical jumping displacement of the middle part, a high-bandwidth, phase-lag-free real-time dynamic wheel load observation model is constructed, and the phase delay defect introduced by the filtering link when the vertical load is indirectly estimated by using the lateral acceleration sensor in the traditional method is effectively avoided.

Description

Technical Field

[0001] This invention relates to the field of corner module technology, and more specifically to an Ackerman adaptive correction system and method for corner modules based on end displacement sensing of transverse leaf springs. Background Technology

[0002] With the development of distributed drive electric vehicle steer-by-wire chassis technology, the corner module architecture, which highly integrates the distributed electric drive assembly, steer-by-wire actuator, brake-by-wire actuator, and suspension system at the wheel ends, has become an important development direction. This architecture eliminates components such as the mechanical steering driveshaft and the through-type lateral stabilizer bar in traditional chassis systems, significantly improving the overall vehicle space utilization and multi-wheel control freedom.

[0003] However, the corner module architecture requires the integrated arrangement of multi-dimensional actuators such as drive, steering, and braking within an extremely limited wheel-end space. Due to the limited wheel-end envelope space, traditional coil spring suspensions are prone to spatial motion interference with rotating components or brake lines under multi-dimensional actuation or large-angle steering conditions, which greatly limits the expansion of the corner module's steering range.

[0004] Furthermore, existing Ackerman steering control systems for modular vehicles mostly employ feedforward calculation methods based on fixed kinematic geometry. That is, the control unit determines the target steering angles of the inner and outer wheels solely based on the real-time input vehicle speed and steering wheel angle, using Ackerman geometric topology formulas.

[0005] However, during actual vehicle operation, significant vertical load transfer inevitably occurs between the inner and outer wheels on the same axle due to lateral acceleration, longitudinal acceleration, or braking deceleration. Since the tire's lateral stiffness characteristics exhibit a markedly nonlinear variation with vertical load, when load transfer during cornering causes the vertical load on the inner and outer tires to deviate significantly from the baseline design, the original Ackermann angle calculated based on fixed geometric relationships will deviate from the ideal optimal tire mechanical response. This can easily lead to nonlinear tire oversaturation on the outer wheel, exacerbating understeer and causing asymmetric abnormal wear on the tires on the same axle. Summary of the Invention

[0006] To address the technical problems existing in the angle module Ackerman control of the prior art, the first aspect of the present invention proposes a technical solution: an angle module Ackerman adaptive correction system based on the end displacement sensing of a transverse leaf spring, comprising: An angle module suspension mechanism includes a lower control arm, a subframe, and a transverse leaf spring as an elastic element. The free end of the transverse leaf spring is connected to the lower control arm through a non-rigid constraint overlapping method, and the free end of the transverse leaf spring has an elastic relative sliding degree of freedom along its longitudinal direction relative to the lower control arm when subjected to vertical load. The displacement sensing unit is used to collect the longitudinal relative sliding displacement ΔX of the free end of the transverse leaf spring relative to the lower control arm and the vertical jump displacement ΔZ of the middle part of the transverse leaf spring relative to the subframe, as a associated deformation signal characterizing the dynamic vertical load change of the wheel. The control unit, electrically connected to the displacement sensing unit, is configured to perform the following steps: Based on the longitudinal relative slip displacement ΔX and the vertical jump displacement ΔZ, the real-time dynamic vertical load F of the corresponding angle module wheel is identified. z ; According to the real-time dynamic vertical load F z Online identification of the real-time tire lateral stiffness difference ΔC between the outer and inner wheels during coaxial turning. α偏差 , where C α Indicates tire lateral stiffness, ΔC α偏差 This indicates the difference in tire lateral stiffness between the outer and inner wheels during coaxial turning. According to the real-time tire lateral stiffness difference ΔC α偏差 The original Ackerman outer wheel target rotation angle δ 外原始 The correction is made to reduce the final target steering angle of the outer wheel, thereby reducing the slip angle of the outer wheel during turning and balancing the tire load utilization of the inner and outer wheels.

[0007] Preferably, the displacement sensing unit includes: The first displacement sensor installed on the lower control arm is used to collect the longitudinal relative sliding displacement ΔX of the free end of the transverse leaf spring relative to the lower control arm; The second displacement sensor installed on the subframe is used to collect the vertical runout displacement ΔZ of the middle part of the transverse leaf spring relative to the subframe. The control unit is equipped with a multi-element decoupling mapping model F. z =G(ΔX,ΔZ), the multivariate decoupling mapping model G is used to perform structural nonlinear decoupling on the longitudinal relative slip displacement ΔX based on the real-time bending state of the transverse leaf spring characterized by the vertical runout displacement ΔZ, so as to calculate the corresponding real-time dynamic vertical load F. z .

[0008] Preferably, the displacement sensing unit is any one of a non-contact laser triangulation rangefinder, a Hall effect linear displacement sensor, or a wire displacement sensor.

[0009] Preferably, it further includes a sensor target fixedly installed on the surface of the free end of the transverse leaf spring, wherein the sensor target and the displacement sensing unit maintain a measurement gap to form a relative motion measurement pair without mechanical rigidity constraint.

[0010] The second aspect of this invention proposes a technical solution: an Ackerman adaptive correction method for an angle module based on the end displacement sensing of a transverse leaf spring. Based on the aforementioned Ackerman adaptive correction system for an angle module based on the end displacement sensing of a transverse leaf spring, the method includes the following steps: Step S1: Use the displacement sensing unit fixedly installed on the lower control arm to collect the longitudinal relative sliding displacement ΔX and vertical jump displacement ΔZ of the free end of the transverse leaf spring; Step S2: Using the control unit, substitute ΔX and ΔZ into the multivariate decoupling mapping model to perform kinematic decoupling, eliminate the influence of the dynamic rotation trajectory of the lower control arm on the slip, and output the real-time vertical load F of each wheel after slip decoupling. zi ; Step S3: Use the control unit to transfer the real-time vertical load F zi The input is fed into a preset tire stiffness observer to identify the real-time lateral stiffness C of the outer wheel during front axle cornering. α外 Real-time lateral stiffness C of the inner wheel during cornering α内 And calculate the difference in lateral stiffness ΔC between the inner and outer wheels of the front axle. α偏差 =C α外 -C α内 ; Step S4: Using the control unit based on the lateral stiffness difference ΔC α偏差 Calculate the outer wheel steering angle correction Δδ 外 And using the outer wheel rotation angle correction amount Δδ 外 The original Ackerman outer wheel target rotation angle δ 外原始 Adjustments are made to reduce the angle and generate the final target turning angle δ. 外目标 And it is sent to the main drive assembly of the corresponding corner module.

[0011] Preferably, in step S2, the test bench calibration and construction method for the multivariate decoupling mapping model G includes the following steps: Bending posture calibration: On the test bench, the lower control arm is driven by the actuator to make vertical jumps throughout the entire stroke. The corresponding vertical jump displacement ΔZ is recorded by the second displacement sensor to construct the bending posture matrix of the leaf spring under the entire stroke. Multidimensional stiffness decoupling calibration: Under constant vertical runout displacement ΔZ at each stage, an increasing vertical load F is applied to the wheel in stages. z And the corresponding longitudinal relative sliding displacement ΔX is recorded by the first displacement sensor; Surface fitting and inverse solution storage: Traverse the discrete points of the vertical runout displacement ΔZ throughout the entire stroke to establish the vertical load F. z The three-dimensional spatial response surfaces of the longitudinal relative sliding displacement ΔX and the vertical jump displacement ΔZ are converted into physical lookup table models and stored in the control unit.

[0012] Preferably, in step S4, the final target rotation angle δ 外目标 The calculation method is as follows: δ 外目标 =δ 外原始 +Δδ 外 Δδ 外 =-K α ·ΔC α偏差 ·f(v x ) Among them, K α The preset adaptive correction gain coefficient, f(v) x The function is a scheduling function based on the longitudinal speed of the vehicle. This scheduling function is configured to converge and decrease the outer wheel steering angle correction Δδ as the vehicle speed increases. 外 The absolute value; the negative sign in the formula indicates that when the lateral stiffness of the outer wheel is greater than that of the inner wheel, the final target steering angle of the outer wheel is reduced, so that the steering posture of the outer wheel shifts towards parallel steering.

[0013] Preferably, the scheduling function f(v) x The expression for ) is: f(v x )=1-v x / v char Among them, v x v represents the real-time longitudinal speed of the vehicle. char The preset vehicle characteristic speed.

[0014] Preferably, between step S1 and step S2, a dynamic protection determination step for unsteady-state impact conditions is further included: The control unit calculates the rate of change of the longitudinal relative slip displacement d(ΔX) / dt at continuous sampling points; when the absolute value of the rate of change |d(ΔX) / dt| is greater than the preset road transient impact threshold, it is determined that the corresponding corner module wheel is under the non-steady-state road high-frequency pulsating impact condition, and the control system jumps into the dynamic transient protection mode. When the absolute value of the rate of change, |d(ΔX) / dt|, is less than or equal to the road transient impact threshold, the vehicle is determined to be in steady-state driving or normal steering condition, and the system normally jumps to execute step S2.

[0015] Preferably, the dynamic transient protection mode includes the following control strategies: Strategy a: Freeze the current outer wheel steering angle correction Δδ using the control unit. 外 This keeps the value at the level it was at the moment before entering the dynamic transient protection mode, in order to suppress the interference of high-frequency transient wheel load fluctuations on the steering target angle planning; Strategy b: Synchronously output active damping control commands to the kingpin drive assembly, and use the reverse electromagnetic torque of the kingpin steering motor to counteract the reverse impact disturbance torque caused by the rebound of the transverse leaf spring and transmitted to the steering knuckle through the lower control arm.

[0016] Compared with the prior art, the advantages of the present invention are as follows: Because the transverse leaf spring of this invention is connected to the ball head at the end of the swing arm, it can achieve a high leverage ratio, significantly reducing the working load of the elastic element and making it suitable for heavy-load vehicles. At the same time, it utilizes the mechanically associated deformation characteristic of longitudinal relative slip displacement caused by the nonlinear elongation of the free end chord length when the transverse leaf spring is bent under load. By integrating this with the vertical jump displacement in the middle, a high-bandwidth, phase-lag-free real-time dynamic wheel load observation model is constructed, effectively avoiding the phase delay defect introduced by the filtering stage when the traditional method of indirectly estimating vertical load using lateral acceleration sensors is used.

[0017] This invention is based on the online observation of the asymmetric lateral stiffness difference of the coaxial tires to achieve adaptive reverse adjustment. When a sharp turning load transfer occurs and the lateral stiffness of the outer tire surges, the target deflection angle of the outer wheel is actively reduced and constrained, causing it to shift towards a parallel steering posture reconstruction. The actual tire lateral angle of the outer wheel is constrained, prompting it to exit the stress saturation nonlinear region. This achieves dynamic balance of load utilization of the coaxial wheels and improves the vehicle's grip limit and anti-skid following stability. Attached Figure Description

[0018] The accompanying drawings are not strictly to scale. In the drawings, each identical or nearly identical component shown in the various figures can be denoted by the same reference numeral. For clarity, not every component is labeled in each figure. Embodiments of various aspects of the invention will now be described by way of example and with reference to the accompanying drawings, wherein: Figure 1 This is a schematic diagram of the Ackerman adaptive correction system for the angle module based on the end displacement sensing of a transverse leaf spring, as shown in this invention. Figure 2 This is a structural schematic diagram of the corner module suspension mechanism shown in this invention; Figure 3 This is a schematic diagram of the longitudinal relative sliding displacement and vertical jump displacement of the transverse leaf spring shown in this invention. Figure 4 This is a schematic diagram of the position of the transverse leaf spring of the wheel under the first load, as shown in this invention. Figure 5 This is a schematic diagram of the position of the transverse leaf spring of the wheel under the second load as shown in this invention; Figure 6 This is a schematic diagram of the steering angle of the inner and outer wheels before modification, as shown in this invention; Figure 7 This is a schematic diagram of the modified steering angle of the inner and outer wheels as shown in this invention; Figure 8 This is a flowchart of the angle module Ackerman adaptive correction method based on end displacement sensing of a transverse leaf spring, as shown in this invention.

[0019] In the diagram: 10, electric drive unit; 20, bogie; 30, braking components; 40, kingpin drive assembly; 50, lower control arm; 51, cover; 60, upper control arm; 70, subframe; 71, bracket; 80, shock absorber; 90, transverse leaf spring; 91, fixed end; 92, middle section; 93, free end; 100, displacement sensing unit; 110, second displacement sensor; 120, first displacement sensor; 200, control unit. Detailed Implementation

[0020] To better understand the technical content of the present invention, specific embodiments are described below in conjunction with the accompanying drawings.

[0021] Example 1 Combination Figure 1 As shown, the first aspect of the present invention proposes a technical solution, an Ackermann adaptive correction system for a corner module based on displacement sensing of the end of a transverse leaf spring, which is applied to a distributed electric drive corner module vehicle, mainly including a corner module suspension mechanism, a displacement sensing unit 100 and a control unit 200.

[0022] Combination Figure 2 As shown, the corner module suspension mechanism includes an electric drive unit 10, a bogie 20, a braking component 30, a kingpin drive assembly 40, a lower control arm 50, an upper control arm 60, a subframe 70, a shock absorber 80, and a transverse leaf spring 90 as an elastic element.

[0023] like Figure 3 As shown, the fixed end 91 of the transverse leaf spring 90 is fixed to the vehicle subframe 70 by a rigid bushing, the middle part 92 of the transverse leaf spring 90 is in an upward arched bending state, and its free end 93 is freely placed on the overlapping surface of the lower control arm 50 in a non-rigid constraint overlapping manner.

[0024] Thus, when the vehicle is subjected to external vertical load, the transverse leaf spring 90 undergoes large bending deformation under vertical compression. When the transverse leaf spring 90 bends, the arch height of its middle part 92 decreases, and the chord length elongates accordingly, thereby driving its free end 93 to generate elastic relative sliding freedom along the longitudinal direction of the vehicle relative to the overlapping surface of the lower control arm 50.

[0025] In order to accurately capture the mechanical deformation signal, the system uses the displacement sensing unit 100 to collect the longitudinal relative sliding displacement ΔX of the free end 93 of the transverse leaf spring 90 relative to the lower control arm 50 and the vertical jump displacement ΔZ of the middle part 92 of the transverse leaf spring 90 relative to the subframe 70, as a associated deformation signal characterizing the change of dynamic vertical load of the wheel.

[0026] Specifically, the displacement sensing unit 100 includes: a first displacement sensor 120 installed on the lower control arm 50, used to collect the longitudinal relative sliding displacement ΔX of the free end 93 of the transverse leaf spring 90 relative to the lower control arm 50; and a second displacement sensor 110 installed on the subframe 70, used to collect the vertical jump displacement ΔZ of the middle part 92 of the transverse leaf spring 90 relative to the subframe 70.

[0027] In an optional embodiment, a cover 51 is constructed above the lower control arm 50. The free end 93 of the transverse leaf spring 90 and the first displacement sensor 120 are both disposed inside the cover 51. A bracket 71 is provided on the subframe 70, and the second displacement sensor 110 is installed inside the bracket 71, thereby avoiding nonlinear pollution and physical damage to the first displacement sensor 120 and the second displacement sensor 110 from external environmental particles such as gravel and mud.

[0028] Optionally, the first displacement sensor 120 and the second displacement sensor 110 can be any one of a non-contact laser triangulation sensor, a Hall effect linear displacement sensor, or a wire-type displacement sensor. Preferably, a laser triangulation sensor is used.

[0029] Specifically, the first displacement sensor 120 of the displacement sensing unit 100 is rigidly fixed on the lower control arm 50, and the measuring end axis of the first displacement sensor 120 is parallel to the longitudinal centerline of the vehicle and faces the free end 93 of the transverse leaf spring 90.

[0030] In an optional embodiment, a sensor target is disposed on the surface of the free end 93 of the transverse leaf spring 90. The target is a metal reflector or a magnetic element, which, in conjunction with a laser displacement sensor or a Hall sensor, forms a non-contact measurement pair.

[0031] The sensor target and the first displacement sensor 120 maintain a measurement gap of 2mm to 5mm.

[0032] The first displacement sensor 120 moves synchronously with the lower control arm 50, and the longitudinal relative sliding displacement ΔX it collects is the absolute amount of sliding of the free end 93 relative to the overlapping surface of the lower control arm 50. However, the longitudinal relative sliding displacement ΔX released by its free end 93 depends not only on the magnitude of the force, but also on the spatial bending curvature of the leaf spring itself.

[0033] It should be understood that when the load changes and the control arm 50 causes the wheel to bounce up and down, the arch height of the middle part 92 of the transverse leaf spring 90 will experience a severe nonlinear drift, resulting in a reconstruction of its vertical stiffness throughout its entire stroke.

[0034] Therefore, in this embodiment, a second displacement sensor 110 is rigidly installed on the subframe 70. The measuring end of the second displacement sensor 110 points vertically downward to the middle 92 of the transverse leaf spring 90, and is used to collect the vertical jump displacement ΔZ of the middle 92 of the transverse leaf spring 90 relative to the subframe 70 in real time. The ΔZ signal is used as a reference feature quantity for the control unit 200 to lock the spatial geometric bending posture of the leaf spring.

[0035] Combination Figure 4 and Figure 5 As shown, when the wheel is subjected to the first load F1, the current vertical displacement collected by the second displacement sensor 110 is Z1, and the current longitudinal relative slip displacement collected by the first displacement sensor 120 is X1. When the wheel is subjected to the first load F1, the current vertical displacement collected by the second displacement sensor 110 is Z2, and the current longitudinal relative slip displacement collected by the first displacement sensor 120 is X2.

[0036] It can be seen that the wheel is subjected to a real-time dynamic vertical load F. z There is a unique, strongly coupled multivariate nonlinear mapping function relationship between the longitudinal relative slip displacement ΔX and the vertical jump displacement ΔZ released in conjunction with it.

[0037] Furthermore, the control unit 200 is equipped with a multi-element decoupling mapping model F. z =G(ΔX,ΔZ), the multivariate decoupling mapping model G is used to characterize and lock the real-time bending state of the transverse leaf spring 90, with the vertical jump displacement ΔZ as a dynamic geometric stiffness anchor. Based on this, the longitudinal relative slip displacement ΔX collected by the first displacement sensor 120 is structurally decoupled nonlinearly, thereby accurately and without delay solving the corresponding real-time dynamic vertical load F in reverse. z .

[0038] In an optional embodiment, the test bench calibration and construction method for the multivariate decoupling mapping model G includes the following steps: On the static calibration bench, the lower control arm 50 is driven by the servo actuator to make vertical displacement within the full stroke. The second displacement sensor 110 records the vertical jump displacement ΔZ of the arch height of the middle 92 of the transverse leaf spring 90 under different wheel jump positions, and constructs the leaf spring bending shape attitude matrix under the full stroke. The vehicle remains stationary under constant vertical runout displacement ΔZ throughout its entire stroke. In this posture, a progressively increasing vertical load F is applied to the diagonal module wheels via the actuator. zAnd the longitudinal relative sliding displacement ΔX corresponding to each load step is recorded by the first displacement sensor 120; Traverse all discrete calibration points of the vertical runout displacement ΔZ throughout the entire stroke to establish the vertical load F. z A three-dimensional spatial response surface relating longitudinal relative slip displacement ΔX and vertical jump displacement ΔZ is constructed. This surface undergoes a nonlinear inverse kinematic transformation, and the F-axis is transformed using a combination of spatial grid lookup and bilinear interpolation algorithms. z =G(ΔX,ΔZ) relational model is stored as a physical lookup table model in the non-volatile memory of control unit 200.

[0039] During actual vehicle operation, the control unit 200 invokes the model with a control cycle of less than 2ms to calculate the real-time dynamic vertical load F after wheel decoupling of the corresponding angle module in real time with high bandwidth. zi (Where i represents the position of the four wheels).

[0040] Furthermore, the control unit 200 calculates the real-time dynamic vertical load F. zi It is then used as a feedforward characteristic signal and input into the built-in tire stiffness observer.

[0041] Optionally, the stiffness observer, based on a semi-empirical Pacejka tire magic formula model, identifies the real-time lateral stiffness C of the outer wheel during current axle cornering online. α外 Real-time lateral stiffness C of the inner wheel during cornering α内 .

[0042] Combination Figure 6 and Figure 7 As shown, traditional control (such as Figure 6 As shown, due to the lack of vertical wheel load feedforward, a load blind zone exists. Under severe cornering conditions, the vehicle load shifts drastically to the outer wheel (right front wheel). At this time, the actual dynamic vertical load on the outer front wheel increases nonlinearly. However, traditional control still maintains the original Ackerman target wheel steering angle δ derived from the original Ackerman geometry. 外原始 This results in an actual sideslip angle δ for the outer wheel (right front wheel) during cornering. 外实际 Excessive force can easily drive the tire force into the nonlinear saturation region, causing a serious imbalance in the load utilization rate of the coaxial tires (defined as the ratio of the actual resultant force of the tire to the current maximum adhesion limit force), exacerbating the understeer of the vehicle and easily triggering understeer.

[0043] To eliminate this deviation in mechanical characteristics, the control unit 200 performs feedforward adaptive correction control on the target steering angle of the Ackermann outer wheel. When the control unit 200 identifies a sharp increase in the lateral stiffness of the outer tire during front axle cornering via the tire stiffness observer online, it calculates the difference in lateral stiffness ΔC between the inner and outer tires of the front axle. α偏差 =Cα外 -C α内 When >0, C α Indicates tire lateral stiffness, ΔC α偏差 This indicates the difference in tire lateral stiffness between the outer and inner wheels during coaxial turning.

[0044] Control unit 200 based on the lateral stiffness difference ΔC α偏差 Substituting the values ​​into the negative feedforward control adaptive correction mechanism, the outer wheel steering angle correction amount Δδ is calculated. 外 : Δδ 外 =-K α ·ΔC α偏差 ·f(v x ); Among them, K α The preset adaptive correction gain coefficient, f(v) x () is a scheduling function based on the longitudinal speed of the vehicle, and its expression is configured as follows: f(v x )=1-v x / v char Among them, v x v represents the real-time longitudinal speed of the vehicle. char This refers to the preset characteristic vehicle speed, v. char The results were obtained through vehicle handling stability bench testing and calibration. Specifically, under a step input condition for steering wheel angle, the difference in lateral stiffness ΔC between the inner and outer wheels of the front axle was measured at different vehicle speeds. α偏差 The curve showing the change in lateral stiffness with vehicle speed is defined as the vehicle speed at which the rate of change of lateral stiffness difference first exceeds a preset threshold. char The calibration value ranges from 60 km / h to 120 km / h. The scheduling function is configured to converge and decrease the outer wheel steering angle correction Δδ as the vehicle speed increases. 外 The absolute value is used to ensure convergent yaw stability when the vehicle is traveling at high speed.

[0045] The ultimate target steering angle δ of the main drive assembly 40 外目标 The calculation method is as follows: δ 外目标 =δ 外原始 +Δδ 外 Because a negative sign is introduced into the feedforward adaptive correction formula, when the cornering load transfer causes the outer wheel's lateral stiffness to be greater than that of the inner wheel (ΔC), α偏差 When the value is greater than 0, the calculated outer wheel angle correction Δδ 外 It is always negative. This means that the control unit 200 is at the geometric Ackerman reference angle δ. 外目标 Based on the geometric reference, the final target deflection offset of the outer wheel was actively reduced and deducted.

[0046] As attached Figure 7 As shown, the final target turning angle δ 外目标 The active reduction in the outer wheel steering angle makes the steering of the outer wheels more inclined towards a parallel steering attitude offset parallel to the vehicle's longitudinal axis. In terms of dynamic response, the active reduction in the outer wheel steering angle adaptively compresses and reduces the actual slip angle of the high-load tires on the outer side of the wheel during cornering, thus reducing its slip state from... Figure 6 The nonlinear saturation edge shown in the diagram safely retreats back into the high linearity design range. The load utilization rate of the outer tire is significantly reduced and tends to dynamically balance with the load utilization rate of the left inner tire, thus rebalancing the left and right lateral force distribution to improve the overall vehicle lateral limit grip and traction performance of the distributed corner module chassis system.

[0047] Example 2 Combination Figure 8 As shown, the second aspect of the present invention proposes a technical solution: an Ackerman adaptive correction method for an angle module based on the displacement sensing of the end of a transverse leaf spring. Based on the aforementioned Ackerman adaptive correction system for an angle module based on the displacement sensing of the end of a transverse leaf spring, the method includes the following steps: Step S1: Use the displacement sensing unit 100 fixedly installed on the lower control arm 50 to collect the longitudinal relative sliding displacement ΔX and vertical jump displacement ΔZ of the free end 93 of the transverse leaf spring 90.

[0048] Specifically, the two sets of signals are synchronously input to the control unit 200 with a sampling period of less than 2ms.

[0049] Furthermore, between step S1 and step S2, the control unit 200 monitors the non-steady-state impact characteristics of the mechanical deformation signal in the time domain in real time to avoid deviation disturbances caused by sudden changes in road surface.

[0050] The control unit 200 utilizes the time-sequential sliding window first-order difference method to calculate the time-domain rate of change d(ΔX) / dt of the longitudinal relative slip displacement in the current sampling period at high frequency. The sliding window length is set to 5-10 consecutive sampling points, the sampling frequency is 500Hz-1000Hz, and the central difference method is used for the difference calculation to improve numerical stability. The control unit 200 compares the absolute value of the rate of change |d(ΔX) / dt| with a preset road transient impact threshold (preferably 120mm / s, adjustable range 80mm / s-150mm / s) in real time. When the absolute value of the rate of change |d(ΔX) / dt| ≤ 120 mm / s, the control unit 200 determines that the vehicle is currently in a steady-state driving condition or a normal quasi-steady-state roll-around cornering condition. The slip signal is determined to be a load characteristic, and the system normally jumps into the normal steady-state feedforward loop. The system normally jumps to execute step S2. When the absolute value of the rate of change |d(ΔX) / dt| > 120 mm / s, the control unit 200 determines that the corresponding wheel angle module is currently experiencing high-frequency pulsating impact from an unsteady road surface. The system immediately disconnects the conventional Ackerman angle correction link, forcibly switches to dynamic transient protection mode, and issues two-layer cross-level active anti-interference defense strategies in parallel: Strategy a: Freeze the current outer wheel steering angle correction Δδ using control unit 200. 外 This keeps the value at the level it was at the moment before entering the dynamic transient protection mode, in order to suppress the interference of high-frequency transient wheel load fluctuations on the steering target angle planning; Strategy b: Synchronously output active damping control commands to the kingpin drive assembly, and use the reverse electromagnetic torque of the kingpin steering motor to counteract the reverse impact disturbance torque caused by the rebound of the transverse leaf spring 90 and transmitted to the steering knuckle through the lower control arm 50.

[0051] It should be understood that when |d(ΔX) / dt| exceeds the threshold, it indicates that the leaf spring is in a non-steady-state stage of high-speed rebound. At this time, the ΔX signal collected by the first displacement sensor 120 contains a large number of high-frequency road excitation components, rather than the actual steady-state load transfer components. If the distorted ΔX is still used for Ackerman correction iteration calculation at this time, it will cause high-frequency chattering of the outer wheel target steering angle, severely deteriorating the driving feel. Compared with using a low-pass filter to filter out noise, this application can immediately cut off the correction link in the first cycle of the impact, without introducing the inherent phase delay of the filtering stage, ensuring that the system can recover to the steady-state correction mode with the fastest response speed after the impact exits.

[0052] Furthermore, the control unit 200 calculates the active reverse damping torque in real time based on the instantaneous amplitude of the rate of change of longitudinal relative slip displacement d(ΔX) / dt. By introducing this active damping term, a negative stiffness suppression effect can be actively introduced on the force chain transmission path before the mechanical impact force is transmitted to the steering knuckle, effectively reducing the physical disturbance of the kingpin angle position caused by high-frequency impacts, thereby protecting the steering actuator from damage caused by hard mechanical impacts.

[0053] Step S2: Using the control unit 200, ΔX and ΔZ are substituted into the multivariate decoupling mapping model for kinematic decoupling, eliminating the influence of the dynamic rotation trajectory of the lower control arm 50 on the slip amount, and outputting the real-time vertical load F of each wheel after slip decoupling. zi (i=1,2,3,4).

[0054] Specifically, because the local stiffness of the transverse leaf spring 90 is constrained by its own spatial curvature during large deformation loads, the longitudinal relative sliding displacement ΔX released by the same vertical force under different vertical jump displacements ΔZ exhibits significant nonlinear differences.

[0055] Therefore, in step S2, by introducing ΔZ to lock the current geometric bending posture of the leaf spring, the absolute amount of pure elastic slip measured by the first displacement sensor 120 is subjected to inverse spatial stiffness mapping. This eliminates the influence of the nonlinear stiffness drift of the leaf spring itself caused by suspension bounce on the wheel load calculation, and can output the real-time dynamic vertical load of each wheel after decoupling with high precision and no delay.

[0056] In an optional embodiment, the test bench calibration and construction method for the multivariate decoupling mapping model G includes the following steps: On the static calibration bench, the lower control arm 50 is driven by the servo actuator to make vertical displacement within the full stroke. The second displacement sensor 110 records the vertical jump displacement ΔZ of the arch height of the middle 92 of the transverse leaf spring 90 under different wheel jump positions, and constructs the leaf spring bending shape attitude matrix under the full stroke. The vehicle remains stationary under constant vertical runout displacement ΔZ throughout its entire stroke. In this posture, a progressively increasing vertical load F is applied to the diagonal module wheels via the actuator. z And the longitudinal relative sliding displacement ΔX corresponding to each load step is recorded by the first displacement sensor 120; Traverse all discrete calibration points of the vertical runout displacement ΔZ throughout the entire stroke to establish the vertical load F. z A three-dimensional spatial response surface relating longitudinal relative slip displacement ΔX and vertical jump displacement ΔZ is constructed. This surface undergoes a nonlinear inverse kinematic transformation, and the F-axis is transformed using a combination of spatial grid lookup and bilinear interpolation algorithms. z =G(ΔX,ΔZ) relational model is stored as a physical lookup table model in the non-volatile memory of control unit 200.

[0057] During system operation, the control unit 200 uses this model to calculate the real-time dynamic vertical load F after the decoupling of the corresponding angle module wheel. zi (Where i represents the position of the four wheels).

[0058] Step S3: Use control unit 200 to transfer the real-time vertical load F zi The input is fed into a preset tire stiffness observer to identify the real-time lateral stiffness C of the outer wheel during front axle cornering. α外 Real-time lateral stiffness C of the inner wheel during cornering α内 And calculate the difference in lateral stiffness ΔC between the inner and outer wheels of the front axle. α偏差 =C α外 -C α内 .

[0059] Step S4: Using the control unit 200 based on the lateral stiffness difference ΔC α偏差 Calculate the outer wheel steering angle correction Δδ 外 And using the outer wheel rotation angle correction amount Δδ 外The original Ackerman outer wheel target rotation angle δ 外原始 Adjustments are made to reduce the angle and generate the final target turning angle δ. 外目标 And it is sent to the main drive assembly of the corresponding corner module.

[0060] Specifically, the control unit 200 identifies and extracts the front axle inner and outer wheel lateral stiffness difference ΔC based on step S3. α偏差 Substitute the values ​​into the feedforward adaptive formula to calculate the outer wheel steering angle correction Δδ. 外 : Δδ 外 =-K α ·ΔC α偏差 ·f(v x ) Among them, K α The preset adaptive correction gain coefficient, f(v) x )=1-v x / v char This is a scheduling function based on the longitudinal speed of the entire vehicle.

[0061] The control unit 200 uses the calculated outer wheel angle correction amount Δδ 外 The original Ackermann outer wheel target rotation angle δ was calculated using Ackermann kinematics. 外原始 Adjustments are made to reduce the amount of friction, and the final target steering angle planning value δ of the outer wheel is generated through composite calculation. 外目标 : δ 外目标 =δ 外原始 +Δδ 外 The control unit 200 will correct the final target angle δ 外目标 The signal is converted into a deflection electrical signal command and sent to the kingpin drive assembly 40 of the corresponding outer wheel angle module. The steering motor in the kingpin drive assembly 40 is precisely positioned to the target correction angle, causing the outer wheel deflection attitude to actively shift towards a parallel steering attitude. By reducing the outer wheel steering angle, the increase in the slip angle of the outer tire during cornering is actively limited and constrained, thereby maintaining the slip response of the high-load tire within the linear optimal force range, rebalancing the left and right lateral grip forces, and making the load utilization rate of the four tires more balanced.

[0062] While the present invention has been disclosed above with reference to preferred embodiments, it is not intended to limit the invention. Those skilled in the art can make various modifications and refinements without departing from the spirit and scope of the invention. Therefore, the scope of protection of the present invention shall be determined by the claims.

Claims

1. An Ackermann adaptive correction system for an angle module based on the end displacement sensing of a transverse leaf spring, characterized in that, include: An angle module suspension mechanism, the angle module suspension mechanism includes a lower control arm (50), a subframe (70) and a transverse leaf spring (90) as an elastic element. The free end (93) of the transverse leaf spring (90) is connected to the lower control arm (50) by a non-rigid constraint overlapping method, and the free end (93) of the transverse leaf spring (90) has an elastic relative sliding degree of freedom along its longitudinal direction relative to the lower control arm (50) when subjected to vertical load. The displacement sensing unit (100) is used to collect the longitudinal relative sliding displacement ΔX of the free end (93) of the transverse leaf spring (90) relative to the lower control arm (50) and the vertical jump displacement ΔZ of the middle part (92) of the transverse leaf spring (90) relative to the subframe (70), so as to serve as the associated deformation signal characterizing the change of dynamic vertical load of the wheel. A control unit (200), electrically connected to the displacement sensing unit (100), is configured to perform the following steps: Based on the longitudinal relative slip displacement ΔX and the vertical jump displacement ΔZ, the real-time dynamic vertical load F of the corresponding angle module wheel is identified. z ; According to the real-time dynamic vertical load F z Online identification of the real-time tire lateral stiffness difference ΔC between the outer and inner wheels during coaxial turning. α偏差 , where C α Indicates tire lateral stiffness, ΔC α偏差 This indicates the difference in tire lateral stiffness between the outer and inner wheels during coaxial turning. According to the real-time tire lateral stiffness difference ΔC α偏差 The original Ackerman outer wheel target rotation angle δ 外原始 The correction is made to reduce the final target steering angle of the outer wheel, thereby reducing the slip angle of the outer wheel during turning and balancing the tire load utilization of the inner and outer wheels.

2. The angle module Ackerman adaptive correction system based on end displacement sensing of a transverse leaf spring according to claim 1, characterized in that, The displacement sensing unit (100) includes: The first displacement sensor (120) installed on the lower control arm (50) is used to collect the longitudinal relative sliding displacement ΔX of the free end (93) of the transverse leaf spring (90) relative to the lower control arm (50); The second displacement sensor (110) installed on the subframe (70) is used to collect the vertical jump displacement ΔZ of the middle part (92) of the transverse leaf spring (90) relative to the subframe (70); The control unit (200) is equipped with a multi-element decoupling mapping model F. z =G(ΔX,ΔZ), the multivariate decoupling mapping model G is used to perform structural nonlinear decoupling on the longitudinal relative slip displacement ΔX based on the real-time bending state of the transverse leaf spring (90) characterized by the vertical jump displacement ΔZ, so as to calculate the corresponding real-time dynamic vertical load F. z .

3. The angle module Ackerman adaptive correction system based on end displacement sensing of a transverse leaf spring according to claim 1, characterized in that, The displacement sensing unit (100) is any one of a non-contact laser triangulation rangefinder, a Hall effect linear displacement sensor, or a wire displacement sensor.

4. The angle module Ackerman adaptive correction system based on end displacement sensing of a transverse leaf spring according to claim 1, characterized in that, It also includes a sensor target fixedly installed on the surface of the free end (93) of the transverse leaf spring (90), and the sensor target and the displacement sensing unit (100) maintain a measurement gap to form a relative motion measurement pair without mechanical rigidity constraint.

5. An Ackerman adaptive correction method for an angle module based on end displacement sensing of a transverse leaf spring, characterized in that, The angular module Ackermann adaptive correction system based on end displacement sensing of a transverse leaf spring, as described in any one of claims 1-4, includes the following steps: Step S1: Use the displacement sensing unit (100) fixedly installed on the lower control arm (50) to collect the longitudinal relative sliding displacement ΔX and vertical jump displacement ΔZ of the free end (93) of the transverse leaf spring (90); Step S2: Using the control unit (200), ΔX and ΔZ are substituted into the multivariate decoupling mapping model for kinematic decoupling, eliminating the influence of the dynamic rotation trajectory of the lower control arm (50) on the slip amount, and outputting the real-time vertical load F of each wheel after slip decoupling. zi ; Step S3: Using the control unit (200), the real-time vertical load F zi The input is fed into a preset tire stiffness observer to identify the real-time lateral stiffness C of the outer wheel during front axle cornering. α外 Real-time lateral stiffness C of the inner wheel during cornering α内 And calculate the difference in lateral stiffness ΔC between the inner and outer wheels of the front axle. α偏差 =C α外 -C α内 ; Step S4: Using the control unit (200) based on the lateral stiffness difference ΔC α偏差 Calculate the outer wheel steering angle correction Δδ 外 And using the outer wheel rotation angle correction amount Δδ 外 The original Ackerman outer wheel target rotation angle δ 外原始 Adjustments are made to reduce the angle and generate the final target turning angle δ. 外目标 And it is sent to the main drive assembly (40) of the corresponding corner module.

6. The angle module Ackerman adaptive correction method based on end displacement sensing of a transverse leaf spring according to claim 5, characterized in that, In step S2, the test bench calibration and construction method for the multivariate decoupling mapping model G includes the following steps: Bending posture calibration: The lower control arm (50) is driven by the actuator to make vertical jump in the whole stroke on the test bench. The corresponding vertical jump displacement ΔZ is recorded by the second displacement sensor (110) to construct the bending posture matrix of the leaf spring in the whole stroke. Multidimensional stiffness decoupling calibration: Under constant vertical runout displacement ΔZ at each stage, an increasing vertical load F is applied to the wheel in stages. z And the corresponding longitudinal relative sliding displacement ΔX is recorded by the first displacement sensor (120); Surface fitting and inverse solution storage: Traverse the discrete points of the vertical runout displacement ΔZ throughout the entire stroke to establish the vertical load F. z The three-dimensional spatial response surfaces of longitudinal relative sliding displacement ΔX and vertical jump displacement ΔZ are converted into physical lookup table models and stored in the control unit (200).

7. The angle module Ackerman adaptive correction method based on end displacement sensing of a transverse leaf spring according to claim 5, characterized in that, In step S4, the final target rotation angle δ 外目标 The calculation method is as follows: d 外目标 =d 外原始 +Dd 外 Dd 外 =-K α ·ΔC α偏差 ·f(v x ) Among them, K α The preset adaptive correction gain coefficient, f(v) x The function is a scheduling function based on the longitudinal speed of the vehicle. This scheduling function is configured to converge and decrease the outer wheel steering angle correction Δδ as the vehicle speed increases. 外 The absolute value; the negative sign in the formula indicates that when the lateral stiffness of the outer wheel is greater than that of the inner wheel, the final target steering angle of the outer wheel is reduced, so that the steering posture of the outer wheel shifts towards parallel steering.

8. The angle module Ackerman adaptive correction method based on end displacement sensing of a transverse leaf spring according to claim 7, characterized in that, The scheduling function f(v) x The expression for ) is: f(v x )=1-v x / v char Among them, v x v represents the real-time longitudinal speed of the vehicle. char The preset vehicle characteristic speed.

9. The angle module Ackerman adaptive correction method based on end displacement sensing of a transverse leaf spring according to claim 5, characterized in that, Between steps S1 and S2, a dynamic protection determination step for unsteady-state impact conditions is also included: The control unit (200) is used to calculate the rate of change d(ΔX) / dt of the longitudinal relative slip displacement ΔX at continuous sampling points; When the absolute value of the rate of change |d(ΔX) / dt| is greater than the preset road transient impact threshold, it is determined that the corresponding corner module wheel is under the non-steady-state road high-frequency pulsating impact condition, and the control system jumps into the dynamic transient protection mode. When the absolute value of the rate of change, |d(ΔX) / dt|, is less than or equal to the road transient impact threshold, the vehicle is determined to be in steady-state driving or normal steering condition, and the system normally jumps to execute step S2.

10. The angle module Ackerman adaptive correction method based on end displacement sensing of a transverse leaf spring according to claim 9, characterized in that, The dynamic transient protection mode includes the following control strategies: Strategy a: Freeze the current outer wheel steering angle correction Δδ using the control unit (200). 外 This keeps the value at the level it was at the moment before entering the dynamic transient protection mode, in order to suppress the interference of high-frequency transient wheel load fluctuations on the steering target angle planning; Strategy b: Synchronously output active damping control commands to the kingpin drive assembly (40), and use the reverse electromagnetic torque of the kingpin steering motor to counteract the reverse impact disturbance torque transmitted to the steering knuckle through the rebound of the transverse leaf spring (90) and the lower control arm (50).

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