A systematic design method for kingpin steering knuckle angle modules

Abstract

The application discloses a systematic design method of kingpin steering knuckle angle module, and belongs to the field of computer-aided design. In order to solve the problem of lacking of systematic process in the existing design method, leading to structural redundancy and low degree of freedom, the application is directed to the kingpin steering knuckle angle module candidate configuration with optimal comprehensive performance, and based on the vehicle performance index, the selection of driving, braking, steering and suspension subsystems and the definition of three-dimensional physical envelope are carried out. Then, the hard point arrangement feasible region is determined through kinematic interference analysis, and the suspension hard point is dynamically optimized by combining a multi-objective optimization algorithm. Finally, virtual integration and comprehensive verification are carried out to obtain the final kingpin steering knuckle angle module. The subjective experience design is changed into objective data driven decision, the unified optimization of kinematics and dynamics is realized, and the integration, reliability and development efficiency of the angle module are significantly improved.

Description

Technical Field

[0001] This invention relates to the field of computer-aided design, and more particularly to a systematic design method for a kingpin steering hub angle module. Background Technology

[0002] With the rapid development of vehicle electrification and intelligence, distributed drive-by-wire chassis platforms have become the mainstream technological trend for new energy vehicles. As an important development direction of distributed chassis, the wheel hub module integrates drive, braking, steering, and suspension systems at the wheel end, supporting independent "drive-brake-steering-suspension" control for each wheel, significantly improving vehicle maneuverability, stability, and environmental adaptability. However, existing wheel hub module technology is still in the exploratory stage, especially lacking a systematic approach in system design, leading to challenges in mass production applications. Currently, wheel hub modules are mainly divided into integral module steering configurations and decoupled kingpin steering configurations based on steering form. Among them, the kingpin steering configuration, due to its unique advantages, is gradually becoming a key path to solve problems such as low control accuracy and response lag.

[0003] In a modular steering configuration, the entire corner module, including drive and braking subsystems, rotates along with the tires during steering. While this configuration simplifies the connection interface, it suffers from drawbacks such as increased rotational inertia, larger vertical space occupation, and excessive bending moments. Furthermore, it's difficult to set wheel-end positioning parameters like kingpin inclination, hindering the expansion of performance limits. In contrast, a decoupled kingpin steering configuration allows only the tires to rotate around the kingpin axis during steering, while the suspension and other supporting structures remain fixed, achieving motion decoupling. This configuration, by optimizing parameters such as kingpin inclination and caster angles, provides natural self-aligning torque, enhancing high-speed stability, while reducing rotational inertia and improving response speed and control precision.

[0004] However, existing design methods for kingpin steering configurations lack a systematic process, and there are contradictions, especially in system parameter matching, hardpoint placement, kinematic interference, and dynamic optimization, which leads to problems such as interference and insufficient performance in practical applications. Summary of the Invention

[0005] In view of the above problems, the present invention provides a systematic design method for a kingpin steering hub angle module, which aims to solve the technical problems of long development cycles caused by fragmented design processes, disconnect between performance and spatial design, and contradictions between kinematic and dynamic optimization in the prior art.

[0006] This invention provides a systematic design method for a kingpin steering hub angle module, comprising: Step S1: For the candidate configuration of the kingpin steering hub angle module with the best overall performance, select and design the steerable subsystem based on the performance requirements of the whole vehicle, and obtain the key performance parameters and three-dimensional physical envelope of each steerable subsystem; Step S2: Based on the key performance parameters and three-dimensional physical envelope of each drive-by-wire subsystem obtained in Step S1, the range of hard point placement without interference is initially determined; the suspension hard point coordinates within the hard point placement range are optimized for multiple objectives: the roll center height optimization objective, the tire alignment parameter variation optimization objective, and the suspension stiffness matching optimization objective are established; sensitivity analysis is performed on the multi-objective optimization to determine the weight coefficients, and the multi-objective optimization is weighted and combined based on the weight coefficients to obtain the final suspension hard point coordinates; Step S3: Perform virtual integration and comprehensive verification of the key performance parameters and three-dimensional physical envelope of each drive-by-wire subsystem in Step S1, as well as the suspension hardpoint coordinates obtained in Step S2, to obtain the final kingpin steering hub angle module.

[0007] Optionally, the method for determining the candidate configuration of the kingpin steering hub angle module with the best overall performance in step S1 is as follows: based on the top-level performance requirements of the whole vehicle, a quantitative analysis and evaluation of multiple candidate configurations is carried out from the aspects of mechanism freedom, motion coupling degree, modularity level and spatial compatibility, and the candidate configuration of the kingpin steering hub angle module with the best overall performance is output.

[0008] Optionally, the drive-by-wire subsystem in step S1 includes a drive motor subsystem, a braking subsystem, a steering subsystem, and a suspension subsystem.

[0009] Optionally, the selection and design of the drive motor subsystem includes: determining the maximum speed cruising power, the climbing torque requirement, and the maximum drive torque of the hub motor based on the vehicle parameters and power performance indicators, and establishing a mapping relationship between torque and radial diameter and axial dimension to determine the three-dimensional physical envelope.

[0010] Optionally, the selection and design of the braking subsystem includes: determining the minimum braking clamping force based on vehicle parameters and braking performance indicators, and matching and selecting the ball screw and electromechanical module to determine the three-dimensional physical envelope.

[0011] Optionally, the design of the steering subsystem includes: establishing a steering load torque calculation model, determining the maximum output torque of the steering subsystem based on steering performance indicators, and determining the three-dimensional physical envelope based on the mapping relationship between torque and the size of the motor and reducer.

[0012] Optionally, the selection and design of the suspension subsystem includes: calculating the equivalent vertical stiffness and equivalent damping of the suspension based on the vehicle parameters and suspension performance indicators, and then determining the required stiffness and load-bearing capacity of the air spring to determine the three-dimensional physical envelope; and designing an active shock absorber based on the maximum working power requirements and anti-roll requirements, and selecting the valve block and electro-hydraulic pump of the kingpin steering hub angle module according to the power and speed requirements of the electro-hydraulic pump.

[0013] Optionally, the method for initially determining the range of interference-free hard point arrangement in step S2 is as follows: Establish a master pin coordinate system to divide the component into rotating and translating components; Using a kinematic model, the dynamic minimum distance between the rotating and translating components is analyzed over the entire range of tire steering angle and wheel bounce displacement. Set a safety gap and record the coordinates of all hard points that meet the interference-free condition to form a preliminary determination of the hard point layout range for interference-free operation.

[0014] Compared with the prior art, the present invention has at least the following beneficial effects: 1. Existing kingpin steering wheel hub corner module design lacks a complete process, resulting in disconnects between subsystem design, layout, and optimization stages, leading to long iteration cycles. This invention proposes a standardized closed-loop process of "subsystem parametric design → hard point multi-objective optimization → integrated simulation verification," making the corner module design process orderly and traceable, improving design effectiveness, and shortening the development cycle.

[0015] 2. In existing systems, the subsystem design of the kingpin steering hub angle module is often carried out in stages with the overall layout, which can easily lead to performance meeting requirements but subsequent integration interference. This invention establishes a mathematical model that directly maps top-level performance requirements to the physical dimensions of key components during the subsystem design phase. This allows for accurate calculation of the three-dimensional envelope of each subsystem in the early stages of design, providing accurate input for subsequent compact integration and avoiding major design rework due to spatial conflicts.

[0016] 3. Existing hard point designs for kingpin steering hub angle modules often compromise on one aspect while pursuing another, either merely satisfying the requirement of no motion interference or only focusing on dynamic performance indicators. This invention innovatively combines kinematic interference analysis based on swept space with dynamic performance optimization based on a multi-objective genetic algorithm. By delineating the deployable region through interference analysis, and then performing dynamic optimization within this region, this method can satisfy both the physical safety requirement of no motion interference under large 90° steering angles and the performance requirements of optimal handling stability and ride comfort, achieving unified optimization of kinematics and dynamics.

[0017] In summary, by applying the design method of this invention, the resulting kingpin steering hub angle module system achieves comprehensive advantages at the product level, including higher integration, better reliability, and faster dynamic response, providing a solid technical guarantee for its mass production application in new energy vehicles. Attached Figure Description

[0018] The accompanying drawings are for illustrative purposes only and are not intended to limit the scope of the invention.

[0019] Figure 1This is a flowchart of the systematic design method of the kingpin steering hub angle module for new energy vehicles according to the present invention; Figure 2 This is a schematic diagram of the kingpin steering hub angle module structure obtained by the present invention; Figure 3 This is a cross-sectional view of the electromechanical brake of the present invention; Figure 4 This is a cross-sectional view of the steering drive assembly of the present invention; Figure 5 This is a front view of the hydraulic active shock absorber of the present invention; Figure 6 This is a cross-sectional view of the hydraulic active shock absorber of the present invention.

[0020] Explanation of reference numerals in the attached figures: 1. Trailing arm, 2. Steering knuckle, 3. Tire, 4. Upper control arm, 5. Lower control arm, 12. External rotor hub motor, 13. Electromechanical brake, 1301. Electromechanical module, 1302. Ball screw, 1303. Parallel shaft gear transmission mechanism, 14. Steering subsystem, 1401. Steering motor, 1402. Harmonic reducer, 1403. Encoder, 15. Hydraulic active shock absorber, 1501. Telescopic air spring, 1502. Active shock absorber, 1503. Valve block, 1504. Electro-hydraulic pump, 15021. Cylinder, 15022. Piston rod, 15023. Blade structure, 16. Kingpin shaft. Detailed Implementation

[0021] To better understand the above-mentioned objectives, features, and advantages of the present invention, the present invention will be further described in detail below with reference to the accompanying drawings and specific embodiments.

[0022] After determining the specific corner module mechanical configuration through a configuration optimization method, this invention proceeds with subsequent subsystem design and selection, kinematic interference analysis, dynamic hardpoint optimization, and system integration verification. See also... Figure 2 The core of the method of the present invention lies in using the kingpin steering hub corner module configuration as a basis, and ensuring the coordinated operation of each subsystem through multi-objective optimization to achieve a compact arrangement and high performance of the corner module.

[0023] A specific embodiment of the present invention, such as Figure 1 A systematic design method for a kingpin steering hub angle module is disclosed, and the specific steps are as follows: Step S1: For the candidate configuration of the kingpin steering hub angle module with the best overall performance, select and design the steerable subsystem based on the vehicle performance indicators, obtain the key performance parameters and three-dimensional physical envelope of each steerable subsystem, and provide preliminary spatial input for subsequent motion interference analysis.

[0024] Furthermore, the steps for determining the candidate configuration of the kingpin steering hub angle module with the best overall performance are as follows: based on the top-level performance requirements of the whole vehicle, a quantitative analysis and evaluation of multiple candidate configurations are conducted from the perspectives of mechanism freedom, motion coupling degree, modularity level and spatial compatibility, and the candidate configuration of the kingpin steering hub angle module with the best overall performance is output.

[0025] Furthermore, the drive-by-wire subsystem includes a drive motor subsystem, a braking subsystem, a steering subsystem, and a suspension subsystem.

[0026] Step S11: Obtain the key performance parameters and three-dimensional physical envelope of the drive motor subsystem of the kingpin steering hub angle module.

[0027] For example, taking the kingpin steering hub angle module obtained in the final design as an example, for the drive motor subsystem, the drive motor subsystem in this example preferably has a compact structure and high transmission efficiency, namely the external rotor hub motor 12.

[0028] Step S111: Obtain the input parameters associated with the drive motor subsystem, including: (1) Vehicle parameters: vehicle weight Full load capacity Tire rolling radius air drag coefficient Windward area ; (2) Performance indicators: maximum speed Maximum gradeability 0km / h to 100km / h acceleration time ; (3) Space constraints of drive motor subsystem layout: The preliminary wheel end space envelope boundary obtained from the candidate configuration of the kingpin steering hub angle module and space analysis is used as the constraint condition for the size of drive motor subsystem.

[0029] Step S112: Based on the input parameters associated with the drive motor subsystems mentioned above, determine the key performance parameters of a single drive motor subsystem through vehicle dynamics calculations, including the maximum cruising power at maximum speed. Torque required for climbing and the maximum drive torque of the hub motor .

[0030] Specifically, the peak power of the drive motor subsystem for new energy vehicles meets the maximum cruising power requirement at the highest vehicle speed. :

[0031] in, For transmission efficiency, an empirical value is used; This is the rolling resistance coefficient; It is the acceleration due to gravity; This refers to air density.

[0032] Specifically, the maximum drive torque of the in-wheel motor in the drive motor subsystem for new energy vehicles. Determined by the maximum gradeability and acceleration performance requirements, the expression is:

[0033] in, This indicates the maximum gradient.

[0034] To meet acceleration performance requirements, the 0km / h to 100km / h acceleration time The expression is:

[0035] in, The vehicle speed corresponding to the inflection point of the constant torque and constant power regions of the drive motor subsystem; To meet the maximum motor torque required for acceleration performance, Vehicle speed during acceleration v h The output torque function of the drive motor after it enters the constant power region.

[0036] Specifically, to meet the requirements for climbing performance and acceleration performance, the maximum drive torque of the hub motor in the drive motor subsystem must satisfy: .

[0037] Step S113: Based on the determined key performance parameters of a single drive motor subsystem, establish a mapping relationship with physical dimensions and preliminarily define the three-dimensional physical envelope of the drive motor subsystem.

[0038] Specifically, the radial diameter of the drive motor subsystem It is related to motor torque and electromagnetic load, and the relationship expression is:

[0039] in, The value is a size factor, which is related to the water / oil cooling method and electromagnetic material grade of the drive motor subsystem, and is obtained by fitting through a database of similar products.

[0040] Specifically, the axial dimension of the drive motor subsystem Related to power density and heat dissipation requirements, the relationship is expressed as follows:

[0041] in, A coefficient related to the material grade and cooling efficiency of the drive motor subsystem; This refers to the rated speed of the hub motor in the drive motor subsystem.

[0042] Furthermore, based on the radial diameter of the drive motor subsystem and the axial dimensions of the drive motor subsystem Determine the three-dimensional physical envelope of the drive motor subsystem.

[0043] Step S12: For the braking subsystem of the kingpin steering hub angle module, taking the vehicle braking performance requirements and layout space as input, through systematic requirement analysis, component matching and multi-objective optimization, a braking subsystem scheme that meets performance requirements and is space-compact is obtained and a three-dimensional physical envelope is obtained.

[0044] For example, the braking subsystem is an electromechanical brake 13.

[0045] Step S121: Obtain the input parameters associated with the braking subsystem, including: (1) Vehicle parameters: vehicle weight Wheelbase L, center of gravity height Distance from center of gravity to front axle Distance from center of mass to rear axle Tire rolling radius Brake disc radius Brake disc friction coefficient Peak road surface adhesion coefficient .

[0046] (2) Braking performance indicators: braking efficiency and response time Continuous and stable load torque wait.

[0047] (3) Spatial constraints of the braking subsystem: The installation space envelope of the braking subsystem is determined based on the overall layout of the aforementioned candidate configurations of the kingpin steering hub angle module, i.e., the maximum allowable length of the braking subsystem. ,width and height .

[0048] The core of the braking subsystem is to provide sufficient braking torque. First, based on the most demanding braking conditions, when the front / rear axle load transfer is at its maximum, the minimum braking clamping force required by each braking subsystem is calculated. .

[0049] Furthermore, taking the front axle as an example, the maximum axle load is calculated. Based on the principle of brake load transfer, the maximum normal reaction force of the ground on the front axle on the peak adhesion surface is... The expression is:

[0050] Furthermore, minimum braking clamping force The expression is:

[0051] in, This represents the maximum normal reaction force exerted by the ground on the rear axle; Specifically, for the front axle, the maximum ground braking force that a single braking subsystem needs to provide. Must meet:

[0052] in, This is the safety factor for braking strength.

[0053] Step S122: Based on the minimum braking clamping force and braking performance indicators, the core actuators of the braking subsystem are matched and selected to form a feasible solution library for ball screws and a feasible solution library for electromechanical modules.

[0054] Specifically, the core actuators of the braking subsystem include ball screws and electromechanical modules.

[0055] Electromechanical module 1301 selection constraints: The electromechanical module provides drive torque, and its selection must meet the output torque and speed requirements to enable it to provide the equivalent load torque under emergency braking and ABS cyclic braking. And reach the required speed. This determines the continuous stall torque of the motor. Peak speed Parameters such as these.

[0056] Ball screw 1302 selection constraints: The ball screw is responsible for converting rotary motion into linear clamping force. Its selection must meet constraints such as static load, dynamic load, critical speed, and bending stability to ensure operation under maximum load. and maximum movement speed To ensure reliable operation, the diameter of the ball screw was determined. , guide l Parameters such as rated load.

[0057] Combination and layout constraint screening: Combine the two feasible solution libraries mentioned above, the ball screw feasible solution library and the electromechanical module feasible solution library, and screen them according to the spatial constraints of the braking subsystem in the overall layout.

[0058] For example, see Figure 3When the braking subsystem is an electromechanical brake 13, the electromechanical brake 13 is arranged in a parallel shaft configuration, and a parallel shaft gear transmission mechanism 1303 is provided between the electromechanical module 1301 and the ball screw 1302, such that its dimensions meet the following constraints:

[0059] in, , For the diameter and length of the ball screw, , For the diameter and length of the electromechanical module, The length of the parallel shaft gear transmission mechanism. The diameter of the parallel shaft transmission mechanism. The lateral layout margin coefficient of the braking subsystem. This is the axial arrangement margin coefficient for the braking subsystem.

[0060] Furthermore, the outer envelope of the braking subsystem is directly determined by the dimensions and arrangement of its selected core actuators (ball screw and electromechanical module), expressed as: .

[0061] Step S13: For the steering subsystem 14 of the kingpin steering hub angle module, obtain its key performance parameters and outer envelope dimensions.

[0062] Further, see Figure 4 The steering subsystem 14 is a steering drive component. The steering subsystem 14 is directly connected to the steering knuckle 2. The steering drive component includes a steering motor 1401, a harmonic reducer 1402, and an encoder 1403. By establishing a steering load torque calculation model, response time analysis, and mapping relationship between key performance parameters and physical dimensions, the scientific design of the steering subsystem is achieved.

[0063] Specifically, the input parameters associated with the steering subsystem 14 are obtained, including: (1) Vehicle parameters: vehicle mass m, wheelbase L, track width B, center of gravity height Distance from center of gravity to front axle Distance from center of mass to rear axle Tire rolling radius Peak road surface adhesion coefficient .

[0064] (2) Steering performance indicators: maximum steering wheel angle (+90°), Maximum response time of the steering subsystem (100ms) Steering control accuracy (±2°), maximum output speed not less than 45° / s.

[0065] (3) Steering subsystem spatial constraints: The installation space envelope of the steering subsystem is determined based on the overall layout of the aforementioned candidate configurations of the kingpin steering hub angle module: that is, the maximum allowable length of the steering subsystem. ,width ,high .

[0066] Furthermore, the expression for the steering subsystem response time is:

[0067] in, The electrical time constant, The electrical time constant, For the stator inductance of the steering motor, For the stator resistance of the steering motor, The equivalent inertia of the steering subsystem. The torque constant of the motor. is the back electromotive force constant.

[0068] Furthermore, the expression for the peak power of the steering motor drive corresponding to the highest output speed of the steering subsystem is:

[0069] in, For steering resistance torque, For steering return torque, To output the highest speed for the steering subsystem, T m This is the equivalent load of the steering subsystem.

[0070] Furthermore, the expression for the steering resistance torque is:

[0071] in, For the maximum load on a single wheel, Tire pressure. For the steering load safety factor, it is generally taken as 1.2. This is the peak adhesion coefficient of the road surface, which is generally taken as 0.7.

[0072] Furthermore, the steering return torque The expression is:

[0073] in, The horizontal distance from the center of the tire to the kingpin. For tire turning angle, The kingpin inclination angle.

[0074] Furthermore, the maximum output torque of the steering subsystem The following relationship must be satisfied:

[0075] Among them, the maximum output torque of the steering subsystem The expression is:

[0076] in, For the steering motor drive torque, For the reduction ratio of the harmonic reducer, This refers to the mechanical transmission efficiency of the harmonic reducer.

[0077] Understandable The length of the steering motor 1401 can be determined through selection and design, directly related to the electromagnetic dimensions of the motor. and diameter , The mechanical dimensions of the harmonic reducer are directly related, and the length of the harmonic reducer can be determined through selection and design. and diameter .

[0078] Furthermore, the outer envelope dimensions of the steering subsystem are:

[0079] in, The encoder 1403 is typically 13mm long. The outer ring diameter of the bearing is typically the diameter of the harmonic reducer. 1.2 times; This indicates the maximum permissible diameter of the steering subsystem.

[0080] Furthermore, the steering subsystem designed in this invention also serves as a suspension guiding mechanism and needs to withstand vertical and radial impact loads from the wheel ends. Based on the determined kingpin steering hub angle module configuration, a topology simulation dynamic model is established on the ADMAS kinematics and dynamics simulation platform. The vertical and radial loads on the steering subsystem under a vehicle speed of 60km / h and a pulse height input of 50mm are extracted as the verification input for the load-bearing performance of the harmonic reducer 1402.

[0081] Step S14: For the suspension subsystem of the kingpin steering hub angle module, based on the overall vehicle performance and space constraints, obtain its key performance parameters and outer envelope dimensions to achieve superior vertical dynamic performance and compact wheel end integration.

[0082] For example, see Figure 5The suspension subsystem is a hydraulic active shock absorber 15. The hydraulic active shock absorber consists of a telescopic air spring 1501, an active shock absorber 1502, a valve block 1503, and an electro-hydraulic pump 1504.

[0083] Step S141: Obtain the input parameters associated with the suspension subsystem, including: (1) Vehicle parameters: Sprout mass Unsprung mass Target frequency offset of suspension subsystem (Typically 1.0-1.5Hz), tire contact point travel Tire static radius .

[0084] (2) Suspension performance indicators: the maximum force used to control the vehicle body attitude Actuator response time System power consumption Active bandwidth (Usually requires ≥15Hz).

[0085] (3) Spatial constraints of the suspension subsystem: The installation space envelope of the suspension subsystem is determined by the overall layout of the aforementioned candidate configurations of the kingpin steering hub angle module: that is, the maximum allowable length of the suspension subsystem. ,width ,high It is necessary to avoid the motion envelope of the steering, drive, and braking systems. It should be noted that the electro-hydraulic pump 1504 and other components are chassis components, and their outer envelope does not take into account the space constraints of the corner module, but the size of the valve block 1503 connected to the active shock absorber 1502 needs to be considered.

[0086] Step S142: Based on the vehicle parameters and suspension performance indicators, calculate the core performance requirements of the suspension subsystem.

[0087] For example, the suspension subsystem is a hydraulic active shock absorber. The core performance requirements of the hydraulic active shock absorber include the equivalent vertical stiffness of the suspension, the equivalent damping of the suspension, the stiffness required by the air spring body, and the load-bearing capacity required by the air spring body.

[0088] Specifically, the equivalent vertical stiffness and equivalent damping requirements of the suspension are determined based on the off-frequency target of the suspension subsystem. and spring mass Calculate the equivalent vertical stiffness of the suspension required at the wheel end. and equivalent damping coefficient The expression is:

[0089] in, The damping ratio is denoted as .

[0090] Furthermore, based on the required equivalent vertical stiffness of the suspension at the wheel end... Design load Leverage ratio (Displacement of hydraulic active shock absorber / wheel bounce displacement) and tilt angle (The angle between the hydraulic active shock absorber and the vertical plane), calculate the required stiffness of the air spring. Bearing capacity Equivalent damping of hydraulic active shock absorbers The expression is: .

[0091] Furthermore, the air spring is a telescopic air spring.

[0092] Step S143: Based on the required stiffness of the air spring and load-bearing capacity By combining the mapping curves of air spring stiffness with pressure, effective area, and volume, a set of effective areas that meet the suspension performance indicators is determined. The initial volume V0 and the working pressure P range are used to determine the outer envelope dimensions of the air spring.

[0093] Specifically, the outer envelope dimensions of the air spring include the air spring diameter and the air spring working height.

[0094] Furthermore, the diameter of the air spring The expression is:

[0095] in, The thickness allowance is 15023 for the air spring's bladder structure.

[0096] Furthermore, the working height of the air spring The expression is:

[0097] in, This is the reference length for the air spring; This indicates the tire's travel distance from the contact point with the ground; in this example, it is -80mm to +100mm.

[0098] Step S144: Design the suspension subsystem based on the maximum working force requirements and anti-roll requirements.

[0099] Specifically, the suspension subsystem design must meet the maximum working force requirements, overcome the inertial force generated by the sprung mass under extreme acceleration, and meet the anti-roll requirements.

[0100] Specifically, see Figure 6The active shock absorber 1502 includes a cylinder 15021, a piston rod 15022, and a bladder structure 15023.

[0101] Furthermore, its maximum working force Must meet:

[0102] in, For maximum vertical control acceleration, For roll stiffness, For the target maximum roll angle, Wheelbase; This indicates the preset load of the air spring.

[0103] Furthermore, considering the fatigue safety factor, the expression for the cylinder inner diameter is:

[0104] in, For the allowable pressure inside the cylinder, Piston rod diameter and cylinder inner diameter The ratio is usually taken as 0.4 to 0.7, and the piston rod diameter needs to be initially selected based on the stability conditions of the pressure rod.

[0105] Based on cylinder inner diameter Selectable shock absorber base length The piston preload stroke is Then the length of the shock absorber for: .

[0106] Step S145: Determine the valve block selection based on the power and speed requirements of the electro-hydraulic pump 1504; determine the electro-hydraulic pump selection based on the response speed of dynamic body control (such as active anti-roll and anti-nose). The electro-hydraulic pump selection must have sufficient power and speed.

[0107] Specifically, the peak power of the electro-hydraulic pump and the maximum lifting speed of the suspension subsystem The expression is:

[0108] in, For the suspension subsystem at the control frequency The target displacement amplitude.

[0109] For valve block selection, the expression for the maximum flow rate through the valve block is:

[0110] Furthermore, according to and valve core allowable flow rate , Typically 5~10 m / s, to obtain the valve core diameter The expression is:

[0111] Furthermore, based on the valve core diameter Select the corresponding standard diameter servo valve.

[0112] Furthermore, the expression for the valve block height is:

[0113] in, The valve body height is positively correlated with the valve core diameter. This is an empirical coefficient. For electrical interface height.

[0114] Furthermore, the key outer envelope dimensions of the suspension subsystem are determined using the above formula. , and .

[0115] .

[0116] Step S2: Based on the key performance parameters and three-dimensional physical envelope of each subsystem determined in Step S1, the range of hard point arrangement without interference is initially determined; the suspension hard point coordinates within the initially determined range of hard point arrangement without interference are optimized by multiple objectives to obtain the suspension hard point coordinates.

[0117] Step S21: Based on the key performance parameters and three-dimensional physical envelope of each wire control subsystem in Step S1, establish the kinematic model and sweep space.

[0118] Based on engineering experience, this invention initially selects the placement positions of the suspension hardpoints relative to the drive motor subsystem, braking subsystem, steering subsystem, and suspension subsystem. By establishing a spatial coordinate system and kinematic model, the interference between the sweep space of rotating components and translational components during steering is quantitatively analyzed, providing accurate constraints for hardpoint optimization.

[0119] Furthermore, the components in the determined kingpin steering hub angle module configuration can be divided into two categories: (1) Rotating components: Components that rotate together with the kingpin axis when the tire rotates, including the outer rotor hub motor 12, electromechanical brake 13, steering subsystem 14, and tire 3; (2) Translational components: Components that do not rotate around the kingpin axis but move with the wheel jump, including the hydraulic active shock absorber 15 and the rigid connecting rod.

[0120] Step S211: Establish a kingpin coordinate system O1-XYZ with the kingpin axis 16 as the kingpin coordinate system. The origin O1 is defined at the intersection of the kingpin axis and the horizontal plane of the tire center. The Z-axis is along the kingpin axis, pointing upwards in the positive direction. The X-axis points directly forward of the vehicle, and the Y-axis points to the left side of the vehicle. The kingpin axis is defined by the kingpin inclination angle. and lean angle α Together, we determine that a unit direction vector can be obtained through coordinate transformation. The expression is: .

[0121] Furthermore, the motion transformation of the component is defined to obtain the motion transformation position of the rotating component; Specifically, the initial position coordinates of any point P on the outer envelope of the rotating component are set as follows: (δ=0, s=0), its position is obtained by first translating and then rotating, resulting in the translated position coordinates of any point P on the rotating component. P s The expression is:

[0122] in, s For wheel jump displacement, Indicates the coordinates of point P after translation; Represents the displacement vector, displacement vector It not only considers the translation of the wheel jump along the Z-axis, but also the small displacements in the X and Y directions that are generated by the suspension kinematics, which can be calculated through the instantaneous center of motion of the suspension.

[0123] Furthermore, select any point on the kingpin axis other than the origin O1. O k Its unit direction vector Then the coordinates of point P after translation P s Rotation about the kingpin axis After the degree, its position is expressed as: .

[0124] Wherein, rotation matrix for:

[0125] Where I is a 3x3 identity matrix, Unit direction vector The cross matrix.

[0126] Understandably, for translational components, their position changes only with wheel hop displacement s, determined by suspension kinematics.

[0127] Step S212: Based on the interference between the rotating component and the translating component during the turning process, obtain the area that can be arranged.

[0128] Specifically, at the tire corner Below, the position of any point on the outer envelope of the rotating component is determined by... Description. When the tire turns... As the target steering range of the tire changes continuously (in this example, [-40°, 90°]), the outer envelope of the rotating component sweeps out a sweeping space; the sweeping space... S Represented as the set of all point trajectories:

[0129] in, Indicates the maximum tire swerve angle on one side of the vehicle; Indicates the maximum tire swerve angle on the other side of the vehicle; Indicates the lower limit of wheel jump displacement; Indicates the upper limit of wheel jump displacement; This is the initial envelope of the rotating component.

[0130] Interference analysis was performed using a discretization method, dividing the tire steering angle and wheel bounce range into N and K equal parts, respectively. δ I , Where I = 1, ..., N; J = 1, ..., K. For each discrete motion position ( δ I , Calculate the straight-line distance between all sampling points on the outer envelope of the rotating and translating components, and record the global minimum distance between the rotating and translating components. The expression is:

[0131] in, This is the initial envelope of the translation component; Represents the tire steering angle at the i-th discrete motion position. and the wheel jump displacement of the Jth discrete motion position The coordinates of any point on the outer envelope surface of the rotating component; Represents the wheel jump displacement at the J-th discrete motion position Under operating conditions, the coordinates of any point on the outer envelope surface of the translation component.

[0132] Step S213: Set safety clearance ε (Usually ≥5mm); If If, then there is no interference; if If interference occurs, adjust the installation position of the rotating component and the position of the translation component's hard point (the suspension hard point, the center of the kinematic pair connecting any two rigid links or between a rigid link and support arm 1), and return to step S212. Record the condition. The set of all possible hard point coordinates This constitutes the feasible domain for hard point placement.

[0133] Furthermore, the feasible region X for hard point placement H The expression is: .

[0134] Step S22: Based on the feasible region X for hard point placement obtained in step S31 H Optimize the coordinates of the suspension hard points.

[0135] Step S221: Taking the optimized double wishbone kingpin steering configuration as an example, the main hard points designed include: the outer point of the upper control arm 4, the inner front point of the upper control arm 4, the inner rear point of the upper control arm 4, the outer point of the lower control arm 5, the inner front point of the lower control arm 5, the inner rear point of the lower control arm 5, the upper mounting point of the shock absorber, and the lower mounting point of the shock absorber. The set of design variables can be expressed as:

[0136] in, These represent the x, y, and z coordinates of the outer point of the upper swing arm under the o-xyz coordinate axis, respectively; These represent the x, y, and z coordinates of the inner front point of the upper swing arm, respectively, along the o-xyz coordinate axis. These represent the x, y, and z coordinates of the inner rear point of the upper swing arm, respectively, along the x-xyz coordinate axis. These represent the x, y, and z coordinates of the outer point of the lower swing arm along the o-xyz coordinate axis. These represent the x, y, and z coordinates of the inner front point of the lower swing arm, respectively, along the x-xyz coordinate axis. These represent the x, y, and z coordinates of the inner rear point of the lower swing arm, respectively, along the x-xyz coordinate axis. These represent the x, y, and z coordinates of the mounting point on the shock absorber, respectively, along the o-xyz coordinate axis. These represent the x, y, and z coordinates of the mounting point of the shock absorber on the o-xyz coordinate axis, respectively.

[0137] It is understandable that the o-xyz coordinate axes are the coordinate axes in the vehicle coordinate system.

[0138] Furthermore, we establish optimization targets for roll center height, tire alignment parameter variation, and suspension stiffness matching.

[0139] Among them, the target for optimizing the roll center height The expression is:

[0140] in, Represents the set of design variables; The height of the suspension roll center. Design a target value for the roll center height, for example, It is 100mm.

[0141] Understandably, the set of design variables Includes the coordinates of all suspension hard points: the x, y, and z axis coordinates of all hard points, including the outer point of the upper control arm, the inner front point, the inner rear point, the outer point of the lower control arm, the upper and lower mounting points of the shock absorber, and the lower mounting point of the shock absorber.

[0142] Among them, the optimization objective for tire alignment parameter changes The expression is:

[0143] in, and These are all weighting coefficients, reflecting the degree of importance attached to the rate of change of tire camber angle and the rate of change of tire toe angle; Indicates the tire camber angle; This indicates the tire toe angle.

[0144] Among them, the suspension stiffness matching optimization objective The expression is:

[0145] in, This is the equivalent stiffness of the suspension calculated based on the suspension hardpoint design variables.

[0146] Step S222: Arrange the feasible region X at the hard point H Generate an initial population X0 and establish a multi-objective optimization problem:

[0147] in, F ( X () is the set of design variables X The multi-objective function vector is optimized by the tilt center height. Optimization goals for changes in tire alignment parameters Optimization target for matching suspension stiffness composition.

[0148] It is understandable that a set of variables is designed here. X The coordinates of the suspension hardpoints are used as design variables.

[0149] Furthermore, the sensitivity of each hard point coordinate to the dynamic performance is determined, expressed as:

[0150] in, For the design variable set X, the first... h Sensitivity of group design variables For the design variable set X, the first... h Group design variables, , These are the standard deviations of the design variables and the objective function, respectively. Indicates the first h The change in the coordinates of the hardpoints of the suspension group; Represents a multi-objective function vector Change in sensitivity Used to assist in determining weighting coefficients .

[0151] Based on the following criteria, the final suspension hardpoint coordinate scheme with the steering subsystem as the suspension guiding mechanism is selected as the preferred scheme, expressed as follows:

[0152] in, The weight coefficients of the p-th optimization objective function in the multi-objective weighted summation are determined based on vehicle performance positioning. Let represent the minimum value that the p-th optimization objective function can achieve; Let p represent the maximum value that the p-th optimization objective function can achieve, where p represents the p-th optimization objective function, and p = 1, 2, 3.

[0153] Step S3: Based on the key performance parameters and three-dimensional physical envelope of each drive-by-wire subsystem obtained in Step S1, and the final suspension hardpoint coordinate scheme obtained in Step S2, this step performs virtual integration and comprehensive verification of the corner module. See [link to relevant documentation]. Figure 5 The feasibility and performance of the design are confirmed through a closed-loop process.

[0154] Step S31: Combining the key performance parameters and three-dimensional physical envelope of each drive-by-wire subsystem determined in Step S1, and the suspension hard point coordinate scheme determined in Step S2, the detailed models of each subsystem are precisely assembled in three-dimensional model building software such as CATIA and Solidworks to form a corner module three-dimensional digital model, which serves as the basis for subsequent simulation verification.

[0155] Step S32: Import the 3D digital model of the corner module into Adams / view multibody dynamics simulation software and model it according to the actual kinematic pairs and constraints. Set a large-angle steering angle in the software, for example, a composite motion condition with a target steering range of -40° to +90° and full wheel hop travel, and perform kinematic simulation. By analyzing the dynamic minimum clearance between the rotating and translating components, quantitatively verify whether there is motion interference, and ensure that the suspension hard point step scheme in step S2 meets the actual motion requirements.

[0156] Step S33: Import the integrated corner module 3D model into the Admas / Car vehicle dynamics simulation platform to construct a vehicle model including the corner module. Verify vehicle handling stability through standard test conditions such as steady-state rotation, angular step, and sinusoidal sweep frequency simulation. Verify vehicle ride comfort through pulse input and random road surface simulation conditions, and obtain key vehicle dynamic indicators. Compare the simulation results with the design goals to verify whether the corner module design contributes to the vehicle's handling stability and ride comfort as expected.

[0157] The above description is only a preferred embodiment of the present invention, but the scope of protection of the present invention is not limited thereto. Any changes or substitutions that can be easily conceived by those skilled in the art within the scope of the technology disclosed in the present invention should be included within the scope of protection of the present invention.

Claims

1. A systematic design method for a kingpin steering hub angle module, characterized in that, include: Step S1: For the candidate configuration of the kingpin steering hub angle module with the best overall performance, select and design the steerable subsystem based on the performance requirements of the whole vehicle, and obtain the key performance parameters and three-dimensional physical envelope of each steerable subsystem; Step S2: Based on the key performance parameters and three-dimensional physical envelope of each drive-by-wire subsystem in Step S1, the range of hard point arrangement without interference is initially determined; the coordinates of the suspension hard points within the range of hard point arrangement are optimized for multiple objectives: the roll center height optimization objective, the tire positioning parameter change optimization objective, and the suspension stiffness matching optimization objective are established. Sensitivity analysis is performed on the multi-objective optimization to determine the weight coefficients. Based on the weight coefficients, the multi-objective optimization is weighted and combined to obtain the final suspension hardpoint coordinates. Step S3: Perform virtual integration and comprehensive verification of the key performance parameters and three-dimensional physical envelope of each drive-by-wire subsystem in Step S1, as well as the suspension hardpoint coordinates obtained in Step S2, to obtain the final kingpin steering hub angle module.

2. The systematic design method according to claim 1, characterized in that, The method for determining the candidate configuration of the kingpin steering hub angle module with the best overall performance in step S1 is as follows: Based on the top-level performance requirements of the whole vehicle, multiple candidate configurations are quantitatively analyzed and evaluated from the perspectives of mechanism freedom, motion coupling degree, modularity level and spatial compatibility, and the candidate configuration of the kingpin steering hub angle module with the best overall performance is output.

3. The systematic design method according to claim 1, characterized in that, The drive-by-wire subsystem mentioned in step S1 includes a drive motor subsystem, a braking subsystem, a steering subsystem, and a suspension subsystem.

4. The systematic design method according to claim 3, characterized in that, The selection and design of the drive motor subsystem includes: determining the maximum cruising power at the highest speed, the torque required for climbing, and the maximum drive torque of the hub motor based on the vehicle parameters and power performance indicators, and establishing a mapping relationship between torque and radial diameter and axial dimension to determine the three-dimensional physical envelope.

5. The systematic design method according to claim 3, characterized in that, The selection and design of the braking subsystem includes: determining the minimum braking clamping force based on vehicle parameters and braking performance indicators, and matching and selecting the ball screw and electromechanical module to determine the three-dimensional physical envelope.

6. The systematic design method according to claim 3, characterized in that, The design of the steering subsystem includes: establishing a steering load torque calculation model, determining the maximum output torque of the steering subsystem based on steering performance indicators, and determining the three-dimensional physical envelope based on the mapping relationship between torque and the size of the motor and reducer.

7. The systematic design method according to claim 3, characterized in that, The selection and design of the suspension subsystem includes: calculating the equivalent vertical stiffness and equivalent damping of the suspension based on the vehicle parameters and suspension performance indicators, and then determining the required stiffness and load-bearing capacity of the air spring to determine the three-dimensional physical envelope; and designing the active shock absorber based on the maximum working power requirements and anti-roll requirements, and selecting the valve block and electro-hydraulic pump of the kingpin steering hub angle module according to the power and speed requirements of the electro-hydraulic pump.

8. The systematic design method according to claim 1, characterized in that, The method for initially determining the range of interference-free hard point arrangement in step S2 is as follows: Establish a master pin coordinate system to divide the component into rotating and translating components; Using a kinematic model, the dynamic minimum distance between the rotating and translating components is analyzed over the entire range of tire steering angle and wheel bounce displacement. Set a safety gap and record the coordinates of all hard points that meet the interference-free condition to form a preliminary determination of the hard point layout range for interference-free operation.

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