Design method of angle-by-wire module for new energy automobile

Abstract

The invention belongs to the technical field of computer aided design, and particularly relates to a drive-by-wire angle module design method for a new energy automobile. According to the design method of the angle-by-wire module, the optimal matching and hierarchical design of a driving subsystem, a braking subsystem, a steering subsystem and a suspension subsystem are respectively carried out by starting from the optimal design of the overall configuration of the angle-by-wire module and taking the overall configuration as the design constraint according to design requirements and arrangement constraints, and the optimal matching and hierarchical design of the driving subsystem, the braking subsystem, the steering subsystem and the suspension subsystem is realized. According to the method, the whole-process design content of the angle-by-wire module is covered, so that a universal and standard angle-by-wire module design method is formed, and the mass production application process of the angle-by-wire module can be accelerated.

Description

Technical Field

[0001] This invention belongs to the field of computer-aided design technology, and in particular relates to a design method for a drive-by-wire corner module for new energy vehicles. Background Technology

[0002] With the rapid development of new energy vehicles, distributed drive-by-wire chassis platforms capable of independent single-wheel drive have become a mainstream technological trend, meeting the demands of vehicle electrification and intelligentization. Based on this, multi-degree-of-freedom controllable, multi-system integrated drive-by-wire corner modules are gradually emerging, revolutionizing traditional automotive chassis architecture. These modules integrate the drive, braking, steering, and suspension systems, positioned at the wheel wells. They are highly modular, enabling independent "drive-braking-steering-active damping" control for each wheel; a single corner module can support the functions of all chassis actuators. Furthermore, the matching of drive-by-wire corner modules with the vehicle allows for three-layer decoupling development of the body, chassis, and actuators. Each corner module acts as a "corner" of the vehicle, supporting plug-and-play integration with the chassis, effectively improving vehicle maneuverability, stability, and environmental adaptability.

[0003] Although various types of steerable cornering modules have been developed for the market, this technology is still in the exploratory stage and requires further standardization and improvement before it can be mass-produced and applied in real vehicles. Currently, designers of steerable cornering modules lack systematic and process-oriented methodological guidance, which fails to fully cover the design boundaries of steerable cornering modules, resulting in significant shortcomings in terms of functional stability and reliability, and design and development cycles. Summary of the Invention

[0004] To address the shortcomings of existing technologies, this invention proposes a design method for steerable angle modules for new energy vehicles. This method addresses design requirements and layout constraints, starting with the optimal design of the overall configuration of the steerable angle module. Using the overall configuration as a design constraint, it conducts optimal matching and hierarchical design of the drive subsystem, braking subsystem, steering subsystem, and suspension subsystem, covering the entire design process of the steerable angle module. This forms a universal and standardized design method for steerable angle modules, which helps to accelerate the mass production and application of steerable angle modules.

[0005] The specific technical solution of the present invention is as follows:

[0006] A design method for a drive-by-wire angle module for new energy vehicles includes the following steps:

[0007] S1: Select the topology of the control-by-wire module configuration based on the application scenarios and target indicators of new energy vehicles;

[0008] S2: Determine the configuration scheme of the drive subsystem, braking subsystem, steering subsystem and suspension subsystem in the steerable angle module, and form the overall layout scheme of the steerable angle module;

[0009] S3: Based on the overall layout scheme of the control-by-wire module, obtain the design boundaries of each subsystem in the control-by-wire module, and complete the design of each subsystem in the control-by-wire module in combination with the performance parameters and functional requirements of the new energy vehicle.

[0010] Preferably, the selection process for the topology of the linear control angle module configuration in S1 specifically includes:

[0011] S1-1: Conduct a full-dimensional topology analysis of the steerable angle module configuration, dividing the steerable angle module configuration into an integral module rotary configuration and a decoupled kingpin steering configuration;

[0012] S1-2: Deep topology of steerable angle module configuration based on the number of vehicle-end interfaces and steering transmission method. The number of vehicle-end interfaces is divided into single-interface connection and multi-directional interface connection; the steering transmission method is divided into direct motor steering and tie rod connection steering.

[0013] S1-3: Eight topological configurations are obtained through S1-1 and S1-2. Taking into account the advantages and disadvantages of each of the eight topological configurations, and considering the application scenarios and target indicators of new energy vehicles, the best topological configuration is selected from the topological configurations.

[0014] Preferably, the overall layout scheme of the line control angle module in S2 is formed in the following way:

[0015] S2-1: Determine the overall layout envelope of the wire-controlled angle module based on the allowable installation space of the optimal topology configuration;

[0016] S2-2: Determine the configuration scheme of the drive subsystem, braking subsystem, steering subsystem and suspension subsystem of the steerable angle module according to the functional requirements of the steerable actuator;

[0017] The drive subsystem uses either a wheel-side motor drive or a wheel hub motor drive.

[0018] The braking subsystem employs an electro-hydraulic brake or an electro-mechanical brake;

[0019] The steering subsystem adopts a direct-drive motor steering unit or a tie-rod connected steering unit.

[0020] The suspension subsystem adopts either passive suspension or active air suspension;

[0021] S2-3: Conduct subsystem decoupling and layout, refine the connection and interaction methods of each subsystem, verify the execution function, and form an overall layout plan for the wire-controlled angle module.

[0022] Preferably, in step S3, the design method of the drive subsystem is as follows:

[0023] Conduct a requirements analysis of the drive subsystem to determine the target power performance indicators of the entire vehicle;

[0024] Perform parameter matching calculations for the drive subsystem to determine the required rated power P of the drive subsystem. drive_e Peak power demand P drive_max Peak speed n drive_max and peak torque demand T drive_max ;

[0025] Based on the required rated power P drive_e Peak power demand P drive_max Peak speed n drive_max and peak torque demand T drive_max Taking into account the overall layout scheme of the wire-controlled angle module, the length L of the drive subsystem's dimensional envelope is... drive Width B drive and high H drive .

[0026] Preferably, in step S3, the design method of the braking subsystem is as follows:

[0027] The braking subsystem employs an electromechanical brake.

[0028] Conduct a requirements analysis of electromechanical brakes, and design the size envelope and required braking clamping force of electromechanical brakes.

[0029] Conduct constraint analysis on electromechanical brakes, solve the feasible solution library for ball screws and the feasible solution library for electromechanical modules separately, and combine them to obtain the matching solution library for electromechanical brakes;

[0030] Conduct constrained analysis of electromechanical brake arrangement, and screen feasible electromechanical brake schemes from the electromechanical brake matching scheme library;

[0031] Multi-objective optimization matching of electromechanical brakes is carried out, with the overall quality, response time and continuous stable load torque as the three indicators of optimization matching. A multi-objective optimization matching design function for electromechanical brakes is established, and a solution is selected from the feasible solution library of electromechanical brakes.

[0032] Preferably, in step S3, the design method of the steering subsystem is as follows:

[0033] The steering subsystem adopts a direct-drive motor steering unit;

[0034] Conduct a demand analysis for direct-drive steering units to determine the required maximum steering torque, required rated speed, and required total power.

[0035] Constraint analysis of the direct-drive steering unit was carried out, and the feasible solution libraries of the reduction mechanism, bolt connection, and steering motor were solved separately. The feasible solution library of the direct-drive steering unit was then obtained by combining them.

[0036] Multi-objective optimization matching of direct-drive steering units is carried out, with overall mass, axial dimension and response speed as the three indicators of optimization matching. A multi-objective optimization matching design function for direct-drive steering units is established, and a solution is selected from the feasible solution library of direct-drive steering units.

[0037] Preferably, in step S3, the design method of the suspension subsystem is as follows:

[0038] The suspension subsystem employs active air suspension.

[0039] Conduct key parameter matching for active air suspension to determine the parameters of air springs and CDC shock absorbers;

[0040] The design of hard point coordinate optimization of active air suspension was carried out. Based on the principle of least squares, the sum of squares of the difference between the designed suspension characteristics and the benchmark suspension characteristics during the simulation was used as the optimization objective. A multi-objective optimization function was established to obtain the spatial arrangement of hard points of active air suspension within the ideal range.

[0041] Finite element verification of active air suspension components under extreme conditions was carried out. Impact loads were calculated under three extreme conditions: off-road, emergency braking, and emergency steering. Based on this, finite element analysis of the active air suspension was performed.

[0042] Compared with the prior art, the beneficial effects of the present invention are as follows:

[0043] 1. The present invention proposes a design method for steerable angle modules for new energy vehicles, involving the entire process of topology analysis and optimization of steerable angle module configuration, overall layout, and detailed design of drive subsystem, braking subsystem, steering subsystem and suspension subsystem. It addresses the design constraints of steerable angle modules and forms a universal and comprehensive steerable angle module design method.

[0044] 2. The present invention proposes a design method for steerable angle modules for new energy vehicles, which systematically describes different configurations, including integral module rotary configuration and decoupled kingpin steering configuration. Based on these two configurations, the advantages, disadvantages and applicable scenarios of the system are analyzed from the perspectives of the number of vehicle-end interfaces and steering transmission method. It covers all existing steerable angle module configurations and provides a systematic reference for the selection of steerable angle module design configurations.

[0045] 3. The present invention proposes a design method for steerable angle modules for new energy vehicles. For the detailed design of the braking system and steering system, it not only considers the constraints of performance parameters such as rated speed, rated torque, and transmission ratio of components such as motor and reduction mechanism, but also calculates functional parameters such as response characteristics and load capacity during operation. Furthermore, it proposes a multi-objective optimization design method that integrates performance parameters and functional parameters, which can select different weights according to the design objectives to obtain the optimal system configuration. Attached Figure Description

[0046] To more clearly illustrate the technical solutions in the embodiments of the present invention or the prior art, the accompanying drawings used in the embodiments will be briefly introduced below. The features and advantages of the present invention can be more clearly understood by referring to the accompanying drawings. The accompanying drawings are schematic and should not be construed as limiting the present invention in any way. For those skilled in the art, other drawings can be obtained based on these drawings without creative effort.

[0047] Figure 1 This is a flowchart of a design method for a drive-by-wire angle module for new energy vehicles according to the present invention.

[0048] Figure 2 This is a flowchart of the preferred topology analysis of the linear control angle module configuration in this invention.

[0049] Figure 3 This is a flowchart illustrating the overall layout scheme of the line control angle module in this invention.

[0050] Figure 4 This is a flowchart illustrating the design of the linear angle control module drive subsystem in this invention.

[0051] Figure 5 This is a design flowchart of the electromechanical brake in the line-controlled angle module braking subsystem of the present invention.

[0052] Figure 6 This is a design flowchart of the motor direct-drive steering unit in the steering subsystem of the line control angle module in this invention.

[0053] Figure 7 This is a flowchart illustrating the design of the active air suspension in the line-controlled angle module suspension subsystem of this invention. Detailed Implementation

[0054] 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. It should be noted that, unless otherwise specified, the embodiments of the present invention and the features thereof can be combined with each other.

[0055] Many specific details are set forth in the following description in order to provide a full understanding of the invention. However, the invention may also be practiced in other ways different from those described herein, and therefore the scope of protection of the invention is not limited to the specific embodiments disclosed below.

[0056] like Figure 1 As shown, the design method for a drive-by-wire angle module for new energy vehicles proposed in this invention includes the following steps:

[0057] S1: Select the topology of the control-by-wire module configuration based on the application scenarios and target indicators of new energy vehicles;

[0058] S2: Determine the configuration scheme of the drive subsystem, braking subsystem, steering subsystem and suspension subsystem in the steerable angle module, and form the overall layout scheme of the steerable angle module;

[0059] S3: Based on the overall layout scheme of the control-by-wire module, obtain the design boundaries of each subsystem in the control-by-wire module, and complete the design of each subsystem in the control-by-wire module in combination with the performance parameters and functional requirements of the new energy vehicle.

[0060] like Figure 2 As shown, the selection process for the topology of the linear control angle module configuration in S1 specifically includes:

[0061] S1-1: Conduct a full-dimensional topology analysis of the steerable angle module configuration, dividing the steerable angle module configuration into an integral module rotary configuration and a decoupled kingpin steering configuration;

[0062] S1-2: Deep topology of steerable corner module configuration based on the number of vehicle-end interfaces and steering transmission method. The number of vehicle-end interfaces can be divided into single-interface connection and multi-directional interface connection; the steering transmission method can be divided into direct motor steering and tie rod connection steering.

[0063] S1-3: Eight topological configurations (topological configuration I to topological configuration IX) can be obtained through S1-1 and S1-2. Considering the advantages and disadvantages of each of the eight topological configurations, and combining the application scenarios and target indicators of new energy vehicles, the best topological configuration is selected from the topological configurations.

[0064] like Figure 3 As shown, the overall layout scheme of the line control angle module in S2 is formed in the following way:

[0065] S2-1: Determine the overall layout envelope of the wire-controlled angle module based on the allowable installation space of the optimal topology configuration;

[0066] S2-2: Determine the configuration scheme of the drive subsystem, braking subsystem, steering subsystem and suspension subsystem of the steerable angle module according to the functional requirements of the steerable actuator;

[0067] The drive subsystem is divided into wheel-side motor drive and wheel hub motor drive. Wheel-side motor drive requires more axial space, but it can act as a sprung mass, effectively reducing unsprung mass and improving vehicle handling stability. Wheel hub motor drive is located inside the wheel rim, effectively utilizing the rim space and occupying less axial space, but the wheel hub motor, as an unsprung mass, is not conducive to vehicle handling stability control.

[0068] The braking subsystem is divided into electro-hydraulic brakes and electro-mechanical brakes. Electro-hydraulic brakes require additional hydraulic lines in the steerable angle module, increasing the complexity of the wiring harness layout. However, because they are located at the vehicle end, they occupy less space within the steerable angle module. Electro-mechanical brakes are fully electric and do not require additional hydraulic lines in the steerable angle module. However, because they are located at the caliper end, they occupy more space within the steerable angle module.

[0069] The steering subsystem is divided into direct-drive motor steering units and linkage-connected steering units. Direct-drive motor steering units offer flexible control and can support 90° steering of the steerable angle module, but they need to be located within the steerable angle module, occupying its compact design space. The linkage-connected steering unit's actuator is located at the vehicle end, not occupying internal space of the steerable angle module, but it cannot achieve 90° steering of the steerable angle module and adds an extra number of interfaces to the vehicle end.

[0070] The suspension subsystem is divided into passive suspension and active air suspension. Passive suspension is simple to arrange and does not require additional active adjustment actuators, but it cannot achieve suspension height adjustment; active air suspension supports suspension height adjustment, but because it has active adjustment actuators, it occupies internal space in the steer-by-wire module.

[0071] S2-3: Conduct subsystem decoupling and layout, refine the connection and interaction methods of each subsystem, verify the execution function, and form an overall layout plan for the wire-controlled angle module.

[0072] like Figures 4-7 S3: Based on the overall layout scheme of the control-by-wire module, obtain the design boundaries of each subsystem in the control-by-wire module, and complete the design of each subsystem in the control-by-wire module in combination with the performance parameters and functional requirements of the new energy vehicle.

[0073] The design methods for each subsystem are as follows:

[0074] 1. Design the drive subsystem

[0075] Parameter matching calculations for drive motor power, speed, and torque are performed, and the selection and design of the drive motor are achieved by comprehensively considering size and weight constraints. Figure 4 As shown, the details are as follows:

[0076] (1) Conduct a drive subsystem requirements analysis to determine the target power performance indicators of the whole vehicle.

[0077] (2) Conduct parameter matching calculations for the drive subsystem, including power, speed and torque requirements.

[0078] a) Based on the target power performance indicators of new energy vehicles, and taking into account three operating conditions—maximum vehicle speed, maximum gradeability, and maximum acceleration—the required rated power of the drive subsystem is determined as follows:

[0079]

[0080] Among them, P vmax P is the power required by the drive subsystem at the highest vehicle speed. slop P is the power required by the drive subsystem under maximum gradeability driving conditions. acc The power required by the drive subsystem under maximum acceleration, v max Where is the maximum vehicle speed, f is the rolling resistance coefficient, M is the total vehicle mass, and C is the maximum speed. D Where A is the air resistance coefficient, A is the frontal area of ​​the new energy vehicle, and v is the air resistance coefficient. slop The required speed for maximum gradeability, α is the gradeability, and t is the speed required for maximum gradeability. m To accelerate time, v m To accelerate t to maximum acceleration m The subsequent vehicle speed, δ is the mass conversion factor.

[0081] This allows us to determine the required rated power P of the drive subsystem. drive_e and peak power demand P drive_max ,as follows:

[0082]

[0083] b) Based on the maximum speed v of new energy vehicles max Determine the peak required speed n of the drive subsystem. drive_max ,as follows:

[0084]

[0085] Among them, R tire This is the tire's rolling radius.

[0086] c) Determine the required peak torque T of the drive subsystem based on the maximum climbing performance requirements. drive_max ,as follows:

[0087]

[0088] Where, α max This represents the maximum gradient.

[0089] (3) Based on the required rated power P drive_e Peak power demand P drive_max Peak speed n drive_max and peak torque demand T drive_max Taking into account the overall layout scheme of the wire-controlled angle module, the length L of the drive subsystem's dimensional envelope is... drive Width B drive and high H drive .

[0090] In summary, when matching and optimizing hub motor drives or wheel-side motor drives, the drive motor with the best output performance should be selected while meeting size constraints and performance requirements.

[0091] 2. Design the braking subsystem.

[0092] For electro-hydraulic brakes, existing conventional hydraulic brake calipers are sufficient. The design of electromechanical brakes (including ball screws and electromechanical modules) is as follows: Figure 5 As shown below:

[0093] (1) Conduct a demand analysis of electromechanical brakes and design the size envelope and required braking clamping force of electromechanical brakes.

[0094] Based on the target design parameters of new energy vehicles and the overall layout scheme of the steerable angle module, the size envelope of the electromechanical brake is determined.

[0095] Based on the load transfer principle during braking, and using the peak road adhesion coefficient... Solve for the normal reaction forces exerted by the ground on the front and rear axles:

[0096]

[0097] Among them, F Zf_max and F Zr_max These are the peak adhesion coefficients of the road surface. The maximum normal reaction force exerted by the ground on the front and rear axles, where L is the front and rear wheelbase. r L is the distance from the center of mass to the rear axle. f h is the distance from the center of mass to the front axle. g The height of the center of mass.

[0098] Therefore, the maximum braking force F that a single electromechanical brake can provide on the ground can be calculated. μ_max ,as follows:

[0099]

[0100] Where, μ p R is the coefficient of friction of the brake disc. discR is the radius of the brake disc. tire F is the tire rolling radius. cl For the required braking clamping force of a single electromechanical brake, α n The braking force safety factor (usually taken as 1.5 to 2).

[0101] F is solved using the following formula. cl :

[0102]

[0103] Among them, F cl_max The maximum required braking clamping force for a single electromechanical brake.

[0104] (2) Conduct constraint analysis of electromechanical brakes, solve the feasible solution library of ball screws and the feasible solution library of electromechanical modules respectively, and combine them to obtain the matching solution library of electromechanical brakes.

[0105] a) Perform constraint calculations for the ball screw, as detailed below:

[0106] Based on the maximum required braking clamping force F of a single electromechanical brake cl_max Solve for the rated static load C a0 Rated dynamic load C a , screw diameter d s With constraints on lead l, the feasible solution library for the ball screw is further solved as follows:

[0107] C a0 ≥α Ca0 F cl_max (10)

[0108]

[0109] Among them, K h The life coefficient of the ball screw (can be determined by...) Calculate, where L h (for the working life of the ball screw), K F It is the load factor of the ball screw (for electromechanical brakes, K). F (1.5), K H K is the coefficient of influence of ball screw stiffness (for electromechanical brakes). H (1), K m The load factor (K) of the ball screw m =F cl_mean That is, F cl_mean K is the average working load force of the ball screw. n For speed coefficient (through) Calculate n m (where n is the average operating speed of the ball screw)s_max α is the maximum speed of the ball screw. Ca0 It is the rated static load safety factor of the ball screw (usually taken as 1 to 3), ω c I is the first-order bending natural frequency of the ball screw. s The minimum sectional moment of inertia of the lead screw in the ball screw (through I) s =π(d s -D p ) 4 (64 calculations), A s The minimum cross-sectional area of ​​the lead screw in the ball screw (via A) s =π(d s -D p ) 2 / 4 calculation), D p The diameter of the ball bearing is... α is the maximum moving speed of the nut in the ball screw. cs λ is the safety factor for the critical speed of the ball screw (generally taken as 0.7 to 1), λ is the support coefficient, and L is the bearing coefficient. screw E is the length of the lead screw in the ball screw. s ρ is the Young's modulus of the lead screw in a ball screw. s D is the screw density. nv Let α be the cycle speed of the leadscrew. f f is the safety factor for bending stress. k δx is the bending coefficient. d The requirement is for precise linear position movement. The accuracy of the rotation angle measured by the sensor.

[0110] b) Perform constraint calculations for the electromechanical modules, as detailed below:

[0111] Based on the maximum required braking clamping force F of a single electromechanical brake cl_max Determine the continuous stall torque T of the electromechanical module based on the motion characteristics of the motion conversion mechanism (ball screw). c Peak speed n p and rated speed constraint n c Based on this, the feasible solution library for the electromechanical module is obtained as follows:

[0112]

[0113] Where, α c T is the safety factor for continuous stall torque. eq_emergency T is the equivalent output load torque of the electromechanical module under emergency braking. eq_ABS The equivalent output load torque of the electromechanical module under ABS braking is given by r, where r is the transmission ratio of the ball screw (via r = 2πi). g / l calculate, i g(η is the transmission ratio of the reduction mechanism, which is 1 here) g η represents the transmission efficiency of the reduction mechanism (100% in this case). s For the transmission efficiency of the ball screw, J m T is the equivalent moment of inertia of the electromechanical brake. m_RMS This represents the root mean square of the motor's output torque. F is the root mean square of the acceleration of the nut in the ball screw. cl_RMS To meet the required braking clamping force F cl The root mean square, for The equivalent value (through) calculate), Let t be the acceleration of the nut moving in the ball screw. d α represents the current cycle operating time of the electromechanical brake. p n is the safety factor for the peak speed of the electromechanical module. e_max This is the maximum load speed of the electromechanical module. α is the maximum output angular velocity of the electromechanical module. nc The safety factor for the rated speed of the electromechanical module. This represents the minimum moving speed of the nut in the ball screw.

[0114] c) Based on the feasible solution library of ball screws and feasible solution library of electromechanical modules obtained by the solution, a matching solution library of electromechanical brakes is obtained by combining them.

[0115] (3) Conduct constrained analysis of the electromechanical brake arrangement, and screen feasible electromechanical brake schemes from the electromechanical brake matching scheme library, as follows:

[0116] Based on the dimensional envelope of the electromechanical brake, and considering both parallel-axis and coaxial arrangements of the ball screw and electromechanical module, the maximum diameter d of the ball screw is calculated. bs_max Maximum length l bs_max and the maximum diameter d of the electromechanical module e_max Maximum length l e_max Based on the constraints, the matching scheme library for electromechanical brakes is screened to obtain a feasible scheme library for electromechanical brakes.

[0117] Coaxial arrangement constraints:

[0118] Parallel axis arrangement constraints:

[0119] Among them, L brake B brake H brake Let α be the length, width, and height of the dimensional envelope of the electromechanical actuator. Bα is the safety factor for the radial dimension of the electromechanical brake (generally taken as 1 to 1.2). L For the axial dimension safety factor of the electromechanical brake (generally taken as 1 to 1.2), l p This refers to the axial length of the parallel gear transmission mechanism in an electromechanical brake arranged in a parallel shaft configuration.

[0120] (4) Conduct multi-objective optimization matching of electromechanical brakes, taking the overall quality, response time and continuous stable load torque as the three indicators of optimization matching, and establish a multi-objective optimization matching design function for electromechanical brakes. According to the objective weights, select the best option from the feasible solution library of electromechanical brakes to achieve the optimal design of comprehensive performance of electromechanical brakes, as follows:

[0121] The total mass m of the electromechanical brake emb The total mass of all its components is as follows:

[0122] m emb =m e +m bs +m p (19)

[0123] Where, m e For the mass of the torque motor, m bs For the ball screw quality, m p For the mass of the parallel gear transmission mechanism, if the electromechanical brake is coaxially arranged, then m p =0.

[0124] Response time t of electromechanical brake ag To allow time for eliminating braking gaps, the electromechanical brake must first ensure that the nut in the ball screw accelerates to its maximum speed. Then at maximum speed Eliminate the remaining braking clearance, t ag The calculation is as follows:

[0125]

[0126] Where, x ag This is the initial braking gap. t1 is the maximum acceleration of the nut in the ball screw, and t1 is the acceleration of the nut in the ball screw to its maximum speed. Time, J e J is the equivalent rotational inertia of the electromechanical module. s J is the equivalent rotational inertia of the lead screw in a ball screw. g J is the equivalent moment of inertia of the parallel gear transmission mechanism (if the electromechanical brake is coaxially arranged, then J) g =0), m n This refers to the mass of the nut in the ball screw.

[0127] Continuous stable load torque T of electromechanical brake stb The equivalent output load torque of the electromechanical module under ABS braking is the same as that shown below:

[0128]

[0129] Based on the above three optimization matching objectives, a multi-objective optimization matching design function for the electromechanical brake is established as follows:

[0130]

[0131] Among them, O emb ω1 is the total mass weighting factor, ω2 is the braking force response time weighting factor, ω3 is the continuous stable braking torque weighting factor, and m is the multi-objective optimization factor for the electromechanical brake (when this optimization factor is calculated to its minimum value, it is considered the optimal matching design). emb _ min For the theoretical minimum mass of the electromechanical brake, t ag_min T is the theoretical minimum braking response time of the electromechanical brake. stb_min This is the theoretical minimum continuous stable load torque for an electromechanical brake.

[0132] 3. Design the steering subsystem

[0133] For tie-rod type steering units, traditional automotive tie-rod type steering units are sufficient. For direct-drive motor steering units, the design is as follows: Figure 6 As shown below:

[0134] (1) Conduct a steering subsystem requirements analysis to determine the required maximum steering torque, required rated speed and required total power;

[0135] Based on the target design parameters of new energy vehicles and the overall layout scheme of the drive-by-wire module, the size envelope of the direct-drive steering unit is determined. Based on the load transfer principle during acceleration and braking, and considering the peak road adhesion coefficient... The maximum normal reaction forces exerted by the ground on the front and rear axles are:

[0136]

[0137] Among them, F dec_Zf_max Peak road surface adhesion coefficient during braking The maximum normal reaction force that the ground can provide to the front axle (the normal reaction force of the ground on the front axle during braking is greater than that on the rear axle), F acc_Zr_max Peak road surface adhesion coefficient during acceleration The maximum normal reaction force that the ground can provide to the rear axle (the normal reaction force of the ground on the rear axle during acceleration is greater than that on the front axle), a acc The absolute value of the maximum acceleration that the vehicle can provide during rapid acceleration, where M is the total vehicle mass, and L is the front and rear wheelbase. r L is the distance from the center of mass to the rear axle. f h is the distance from the center of mass to the front axle. g The height of the center of mass.

[0138] Therefore, the maximum vertical load F experienced by any axle of a new energy vehicle under various operating conditions can be obtained. z_max ,as follows:

[0139] F z_max =max(F dec_Zf_max ,F acc_Zr_max )(twenty four)

[0140] Since the steering subsystem in the steerable angle module is an independent steering system, with each wheel driven by an independent direct-drive motor steering unit, the steering resistance torque of each wheel on the front and rear axles needs to be analyzed when designing the steering subsystem parameters. Considering that the direct-drive motor steering units used on all wheels of the front and rear axles are identical, and taking into account the axle load transfer effect, the maximum required steering torque T that a single direct-drive motor steering unit can output is calculated. max That is, the maximum steering resistance torque T that the direct-drive steering unit needs to overcome during operation. w_max ,as follows:

[0141]

[0142] Where μ is the coefficient of sliding friction between the tire and the road surface, p min α is the minimum tire pressure required for normal vehicle operation. tw This is the safety factor for steering torque.

[0143] Secondly, based on the target design parameters of the steering subsystem, the required rated speed ω that the direct-drive steering unit can output is calculated. tw for:

[0144]

[0145] Where, δ max The maximum wheel angle achievable by the steering subsystem in normal driving mode, t res The maximum response time of the steering subsystem while ensuring vehicle handling performance.

[0146] Therefore, the total required power P that a single motor direct-drive steering unit can output is calculated as follows:

[0147] P = Tmax ω tw (27)

[0148] (2) Conduct constraint analysis of the direct-drive steering unit, solve the feasible solution library of the deceleration mechanism, the feasible solution library of the bolt connection and the feasible solution library of the steering motor respectively, and combine them to obtain the feasible solution library of the direct-drive steering unit.

[0149] a) Perform constraint calculations for the deceleration mechanism, as follows:

[0150] Based on the maximum vertical load F on any axle of a new energy vehicle under various operating conditions z_max Solve for the maximum axial load F of the reduction mechanism of the direct-drive steering unit. a Maximum radial load F r The constraints on the maximum allowable bending moment M are as follows:

[0151]

[0152] Among them, F r_side (F r_side =G α F z_max ), F r_brk (F r_brk =μF z_max G represents the lateral force and braking force exerted on the tires by the ground, respectively. α denoted as the linear coefficient relating the vertical and lateral forces of the tire, μ is the coefficient of sliding friction between the tire and the road surface, and h is the distance from the ground to the center of the longitudinal height of the deceleration mechanism.

[0153] b) Perform constraint calculations for the steering connection flange, as follows:

[0154] Based on the maximum axial load F of the direct-drive steering unit reduction mechanism a Maximum radial load F r Given the maximum allowable bending moment M, solve for the feasible bolt connection scheme library for the turning flange, including the number of bolts n and the bolt type (which needs to be selected by referring to the table), as follows:

[0155]

[0156] in, These are the design values ​​of the compressive bearing capacity and shear bearing capacity of a single bolt, respectively, in N. vi r i Let d be the shear force borne by the i-th bolt on the steering connection flange and the distance from the axis of the deceleration mechanism, respectively (i = 1, 2, 3…n). trWhere n is the bolt rod diameter, ∑t is the total thickness of the connected parts resisting compression in the same direction (here, it is a single shear connection between the steering connecting plate and the steering connecting flange, taking the thickness of the thinner plate between the steering connecting plate and the steering connecting flange), and n is the diameter of the bolt rod. v This represents the number of shear surfaces of the bolt. These are the design values ​​for the bolt's compressive strength and shear strength, respectively.

[0157] The formula for calculating allowable bending moment constraints can be simplified as follows:

[0158] N v1 / r1=N v2 / r2=N v3 / r3=…=N vn / r n (30)

[0159] Therefore, the allowable bending moment constraint can be expressed as:

[0160]

[0161] c) Perform the steering motor constraint calculations as follows:

[0162] Based on the maximum steering torque T that a single motor direct-drive steering unit can output. max The rated speed ω that the direct-drive steering unit can output. tw Given the total power P that a single direct-drive steering unit can output, the feasible solution library for the steering motor is obtained as follows:

[0163]

[0164] Among them, T motor To achieve the maximum torque required by the steering motor, n motor The required rated speed (i) for the steering motor t_total =i t i r ), P motor To meet the rated power requirements of the steering motor, i t_total i t and i r These represent the total transmission ratio of the direct-drive steering unit, the total transmission ratio of the reversing drive, and the total transmission ratio of the reduction mechanism, respectively. t_total η e η t η r These are the total efficiency of the direct-drive steering unit, the total efficiency of the steering motor, the total efficiency of the reversing transmission, and the total efficiency of the reduction mechanism (η). t_total =η e η t η r ).

[0165] Note that the aforementioned reversing transmission is a mechanism that changes the transmission direction between the steering motor and the reduction mechanism, thus facilitating the arrangement of the steerable angle module, such as a right-angle steering gear, a coaxial drive steering gear, etc. If this device is not available, then... t =1 and η t =1.

[0166] d) Based on the feasible solution library of the deceleration mechanism, the feasible solution library of the bolt connection, and the feasible solution library of the steering motor, a feasible solution library of the direct-drive steering unit is obtained by combining them.

[0167] (3) Conduct multi-objective optimization matching of the direct-drive steering unit. Take the overall mass, axial dimension and response speed as the three indicators of optimization matching, establish the multi-objective optimization matching design function of the direct-drive steering unit, and select the best option from the feasible scheme library of the direct-drive steering unit according to the target weight, so as to achieve the optimal design of the comprehensive performance of the direct-drive steering unit.

[0168] The total mass m of the direct-drive steering unit tw The total mass of all its components is as follows:

[0169] m tw =m motor +m t +m r +m chasf +m outf (33)

[0170] Where, m motor For the mass of the steering motor, m t For the mass of the commutator, m r For the mass of the reduction mechanism, m chasf For the quality of the steering connection flange, m outf To determine the output flange mass, if the reduction mechanism uses shaft output instead of flange output, then m outf =0.

[0171] Axial dimension l of direct-drive motor steering unit tw The distance from the bottom of the output flange to the bottom of the swivel connection flange is as follows:

[0172] l tw =l r +l outf -Δl chasf (34)

[0173] Among them, l r l is the axial length of the reduction mechanism. outf The axial length of the output flange is given. If the reduction mechanism uses shaft output instead of flange output, then l outf =0, Δl chasfThis is the distance from the bottom of the steering connection flange to the top of the deceleration mechanism. This value is positive when the bottom of the steering connection flange is lower than the top of the deceleration mechanism.

[0174] The response time τ of a direct-drive steering unit can be considered as the electrical time constant τ of the steering motor. e Mechanical time constant τ m Time τ for eliminating transmission backlash with the reduction mechanism ag The sum is calculated as follows:

[0175] τ=τ e +τ m +τ ag (35)

[0176] Based on the above three optimization matching objectives, a multi-objective optimization matching design function for the direct-drive steering unit is established as follows:

[0177]

[0178] Among them, O tw ω is the multi-objective optimization factor for the direct-drive steering unit (when this optimization factor is calculated to its minimum value, it is considered the optimal matching design). tw1 ω is the overall quality weighting factor for the steering unit. tw2 ω is the axial dimension weighting factor for the steering unit. tw3 m is the weighting factor for the steering unit response time. tw_min For the theoretical minimum mass of the steering unit, l tw_min τ is the theoretical minimum axial dimension of the steering unit. min This represents the theoretical minimum response time of the steering unit.

[0179] 4. Design the suspension subsystem.

[0180] For passive suspension, the traditional passive suspension design of automobiles is sufficient. For active air suspension design, such as... Figure 7 As shown below:

[0181] (1) Conduct key parameter matching of active air suspension to determine the parameters of air spring and CDC shock absorber.

[0182] a) Conduct key parameter matching for the air spring, including the air spring stiffness K. sus and suspension offset frequency n sus ,as follows:

[0183] Air spring stiffness K sus Directly affects the elastic force F of the air spring sus Vertical travel x of the air spring sus The relationship is calculated as follows:

[0184]

[0185] Among them, A e P is the effective working area of ​​the air spring. e f is the effective pressure during the operation of the air spring. sus P is the gas polytropic index, P0 is the initial pressure of the air spring, and P a Let V be the atmospheric pressure, V0 be the initial effective volume of the air spring, and V e This refers to the effective volume during the operation of the air spring.

[0186] A e P e and V e All are quantities of change, and the calculation formulas are as follows:

[0187]

[0188] Where A0 is the initial effective working area of ​​the air spring, υ sus ρ is the rate of change of the effective working area of ​​the air spring. sus This represents the rate of change of the effective volume within the air spring airbag.

[0189] When the vertical travel x of the air spring sus When P is 0, e =P0 and V e =V0, then the net stiffness K of the air spring in equilibrium state is... sus_0 for:

[0190]

[0191] Suspension frequency n is calculated based on suspension stiffness. sus ,as follows:

[0192]

[0193] Where, m sp For the single-wheel sprung mass of the entire vehicle, K sus_0 Substituting the values, we can obtain the air spring deflection frequency n in static equilibrium. sus_0 ,as follows:

[0194]

[0195] b) Conduct key parameter matching for the CDC vibration damper, including the damping coefficient C of the CDC vibration damper. cdc Maximum unloading force F d_max and its working cylinder diameter D cdc ,as follows:

[0196] The relative damping coefficient ψ of an active air suspension is defined as:

[0197]

[0198] The range of values ​​for the relative damping coefficient ψ of the active air suspension is determined based on the vehicle design constraints, i.e., [ψ min ,ψ max ].

[0199] The damping coefficient C of the CDC shock absorber can then be calculated. cdc The range of values ​​for is as follows:

[0200] 4πψ min m sp n sus ≤C cdc ≤4πψ max m sp n sus (43)

[0201] Meanwhile, to avoid damage to the CDC vibration damper from excessive instantaneous impact force, it is necessary to design the maximum unloading force F of the CDC vibration damper. d_max ,as follows:

[0202] F d_max =C r_max v w (44)

[0203] Among them, C r_max v is the maximum tensile damping coefficient of the CDC damper. w The piston unloading speed of a CDC shock absorber is typically calculated as follows:

[0204] v w =2πn sus A w (45)

[0205] Among them, A w The distance traveled by the wheel.

[0206] Therefore, the working cylinder diameter D of the CDC shock absorber can be obtained. cdc ,as follows:

[0207]

[0208] Where [p] is the allowable pressure inside the CDC shock absorber cylinder, and λ d It is the ratio of the piston rod diameter to the cylinder diameter.

[0209] (2) Conduct optimization design of hard point coordinates of active air suspension. Based on the principle of least squares, take the sum of squares of the difference between the designed suspension characteristics and the benchmark suspension characteristics during the simulation as the optimization objective, establish a multi-objective optimization function, and obtain the spatial arrangement of hard points of active air suspension within the ideal range.

[0210] The mathematical description of the active air suspension hardpoint optimization problem is as follows:

[0211]

[0212] F(x) is a multi-objective optimization function, where m is taken as 4. f1(x), f2(x), f3(x), and f4(x) are the optimization factors for wheel camber angle, wheel toe angle, roll center height, and track width, respectively. The optimization factor f... i (x) is represented as:

[0213] f i (x)=|f KO_i -f KA_i (x)| (48)

[0214] Among them, f KO_i f represents the ideal value of the optimization factor. KA_i (x) represents the actual value of the optimization factor.

[0215] In the optimization function, x = (x1, x2, ..., x...) n () represents the selected hard point coordinate parameters, and x is set. 1min ≤x1≤x 1max (x i_min With x i_max They represent x respectively i (Minimum and maximum values). Taking a double wishbone suspension as an example, n represents the number of hard points, which is 24. These correspond to 24 coordinate parameters: the X, Y, and Z coordinates of the lower control arm front point, lower control arm rear point, lower control arm outer point, upper control arm front point, upper control arm rear point, upper control arm outer point, upper kingpin upper end point, and kingpin lower end point. Different coordinate parameters correspond to different objective function values.

[0216] As can be seen from the mathematical description of the active air suspension hardpoint optimization problem, the optimization process involves multiple objective variables, and these variables conflict and influence each other, making it a typical multi-objective optimization problem. To select suitable design variables, sensitivity analysis of the objective function is necessary. Therefore, a large number of design variables that may affect the objective function need to be selected as analysis factors to analyze their contribution to changes in the objective function. The design variables with larger contributions, i.e., those with higher sensitivity, are selected. This reduces the computational load in the optimization process while obtaining more ideal optimization results.

[0217] (3) Conduct finite element verification of active air suspension components under extreme conditions. Calculate the impact load under three extreme conditions: off-road condition, emergency braking condition, and emergency steering condition. Based on this, perform finite element analysis of the active air suspension.

[0218] a) Impact load calculations are performed for three extreme operating conditions: off-road, emergency braking, and emergency steering, as detailed below:

[0219] The first extreme operating condition is off-road. When the vehicle crosses an obstacle, due to the height difference, the active air suspension will be subjected to an impact force from the direction of gravity of the road surface. Ignoring the influence of suspension stiffness, the impact load of the dent is calculated as follows:

[0220]

[0221] Among them, F Z_max k represents the maximum vertical load at the contact point between the wheel and the ground of a single steerable angle module. dl Here, k is the dynamic load coefficient (usually taken as 2.5), and g is the acceleration due to gravity.

[0222] The second extreme operating condition is emergency braking. During emergency braking, the loads on the front and rear axles of the vehicle change. To simplify the calculation model, only the forces in the X and Z directions are considered, while the force in the Y direction is ignored. The forces on the left and right front wheels are treated as the same during braking. During emergency braking, the vehicle load will transfer to the front axle. The force at the wheel is calculated using a braking acceleration of 0.9g, and the front axle load transfer ΔF is used. Z(0.9g)_brake for:

[0223]

[0224] Therefore, the vertical force F of a single front wheel after the front axle load transfer Zf_brake for:

[0225]

[0226] The third extreme operating condition is the emergency steering condition. During an emergency steering maneuver, the tire load will shift, and the outer wheel will bear a greater vertical force. To simplify the calculation model, only the forces in the Y and Z directions of the vehicle are considered, ignoring the forces in the X direction. The forces on the left and right front wheels during braking are considered to be the same. During the emergency steering maneuver, the vehicle load will shift towards the front axle. Here, a lateral acceleration of 0.9g is used to calculate the force at the wheel. The front axle load transfer ΔF under this condition is... Z(0.9g)_steer for:

[0227]

[0228] Where B is the wheel track of the entire vehicle.

[0229] The vertical force on the outer front wheel after the front axle load transfer is F. Zf_steer :

[0230]

[0231] b) Apply loads and mesh the active air suspension components. Apply the calculated impact loads under extreme conditions to the active air suspension geometric model, and discretize the active air suspension geometric model into finite elements through mesh generation, and perform finite element analysis.

[0232] In this invention, unless otherwise explicitly specified and limited, the terms "installation," "connection," "linking," and "fixing," etc., should be interpreted broadly. For example, they can refer to a fixed connection, a detachable connection, or an integral part; they can refer to a mechanical connection or an electrical connection; they can refer to a direct connection or an indirect connection through an intermediate medium; they can refer to the internal communication of two components or the interaction between two components. Those skilled in the art can understand the specific meaning of the above terms in this invention according to the specific circumstances.

[0233] In this invention, unless otherwise explicitly specified and limited, "above" or "below" the second feature can include direct contact between the first and second features, or contact between the first and second features through another feature between them. Furthermore, "above," "over," and "on top" of the second feature includes the first feature directly above or diagonally above the second feature, or simply indicates that the first feature is at a higher horizontal level than the second feature. "Below," "below," and "under" the second feature includes the first feature directly below or diagonally below the second feature, or simply indicates that the first feature is at a lower horizontal level than the second feature.

[0234] In this invention, the terms "first," "second," "third," and "fourth" are used for descriptive purposes only and should not be construed as indicating or implying relative importance. The term "multiple" refers to two or more unless otherwise expressly defined.

[0235] The above description is merely a preferred embodiment of the present invention and is not intended to limit the invention. Various modifications and variations can be made to the present invention by those skilled in the art. Any modifications, equivalent substitutions, improvements, etc., made within the spirit and principles of the present invention should be included within the scope of protection of the present invention.

Claims

1. A design method for a drive-by-wire angle module for new energy vehicles, characterized in that, Includes the following steps: S1: Select the topology of the control-by-wire module configuration based on the application scenarios and target indicators of new energy vehicles; S2: Determine the configuration scheme of the drive subsystem, braking subsystem, steering subsystem and suspension subsystem in the steerable angle module, and form the overall layout scheme of the steerable angle module; S3: Based on the overall layout scheme of the control-by-wire module, obtain the design boundaries of each subsystem in the control-by-wire module, and complete the design of each subsystem in the control-by-wire module in combination with the performance parameters and functional requirements of the new energy vehicle.

2. The design method for a drive-by-wire angle module for new energy vehicles according to claim 1, characterized in that, The selection process for the topology of the line control angle module configuration in S1 specifically includes: S1-1: Conduct a full-dimensional topology analysis of the steerable angle module configuration, dividing the steerable angle module configuration into an integral module rotary configuration and a decoupled kingpin steering configuration; S1-2: Deep topology of steerable angle module configuration based on the number of vehicle-end interfaces and steering transmission method. The number of vehicle-end interfaces is divided into single-interface connection and multi-directional interface connection; the steering transmission method is divided into direct motor steering and tie rod connection steering. S1-3: Eight topological configurations are obtained through S1-1 and S1-2. Taking into account the advantages and disadvantages of each of the eight topological configurations, and considering the application scenarios and target indicators of new energy vehicles, the best topological configuration is selected from the topological configurations.

3. The design method for a drive-by-wire angle module for new energy vehicles according to claim 2, characterized in that, The overall layout scheme of the S2 centerline control angle module is formed in the following way: S2-1: Determine the overall layout envelope of the wire-controlled angle module based on the allowable installation space of the optimal topology configuration; S2-2: Determine the configuration scheme of the drive subsystem, braking subsystem, steering subsystem and suspension subsystem of the steerable angle module based on the functional requirements of the steerable actuator; The drive subsystem uses either a wheel-side motor drive or a wheel hub motor drive. The braking subsystem employs an electro-hydraulic brake or an electro-mechanical brake; The steering subsystem adopts a direct-drive motor steering unit or a tie-rod connected steering unit. The suspension subsystem adopts either passive suspension or active air suspension; S2-3: Conduct subsystem decoupling and layout, refine the connection and interaction methods of each subsystem, verify the execution function, and form an overall layout plan for the wire-controlled angle module.

4. The design method for a drive-by-wire angle module for new energy vehicles according to claim 3, characterized in that, In S3, the design method of the drive subsystem is as follows: Conduct a requirements analysis of the drive subsystem to determine the target power performance indicators of the entire vehicle; Perform parameter matching calculations for the drive subsystem to determine the required rated power P of the drive subsystem. drive_e Peak power demand P drive_max Peak speed n drive_max and peak torque demand T drive_max ; Based on the required rated power P drive_e Peak power demand P drive_max Peak speed n drive_max and peak torque demand T drive_max Taking into account the overall layout scheme of the wire-controlled angle module, the length L of the drive subsystem's dimensional envelope is... drive Width B drive and high H drive .

5. The design method for a drive-by-wire angle module for new energy vehicles according to claim 4, characterized in that, In S3, the design method of the braking subsystem is as follows: The braking subsystem employs an electromechanical brake. Conduct a requirements analysis of electromechanical brakes, and design the size envelope and required braking clamping force of electromechanical brakes. Conduct constraint analysis on electromechanical brakes, solve the feasible solution library for ball screws and the feasible solution library for electromechanical modules separately, and combine them to obtain the matching solution library for electromechanical brakes; Conduct constrained analysis of electromechanical brake arrangement, and screen feasible electromechanical brake schemes from the electromechanical brake matching scheme library; Multi-objective optimization matching of electromechanical brakes is carried out, with the overall quality, response time and continuous stable load torque as the three indicators of optimization matching. A multi-objective optimization matching design function for electromechanical brakes is established, and a solution is selected from the feasible solution library of electromechanical brakes.

6. The design method for a drive-by-wire angle module for new energy vehicles according to claim 5, characterized in that, In S3, the design method of the steering subsystem is as follows: The steering subsystem adopts a direct-drive motor steering unit; Conduct a demand analysis for direct-drive steering units to determine the required maximum steering torque, required rated speed, and required total power. Constraint analysis of the direct-drive steering unit was carried out, and the feasible solution libraries of the reduction mechanism, bolt connection, and steering motor were solved separately. The feasible solution library of the direct-drive steering unit was then obtained by combining them. Multi-objective optimization matching of direct-drive steering units is carried out, with overall mass, axial dimension and response speed as the three indicators of optimization matching. A multi-objective optimization matching design function for direct-drive steering units is established, and a solution is selected from the feasible solution library of direct-drive steering units.

7. The design method for a drive-by-wire angle module for new energy vehicles according to claim 6, characterized in that, In S3, the design method of the suspension subsystem is as follows: The suspension subsystem employs active air suspension. Conduct key parameter matching for active air suspension to determine the parameters of air springs and CDC shock absorbers; The design of hard point coordinate optimization of active air suspension was carried out. Based on the principle of least squares, the sum of squares of the difference between the designed suspension characteristics and the benchmark suspension characteristics during the simulation was used as the optimization objective. A multi-objective optimization function was established to obtain the spatial arrangement of hard points of active air suspension within the ideal range. Finite element verification of active air suspension components under extreme conditions was carried out. Impact loads were calculated under three extreme conditions: off-road, emergency braking, and emergency steering. Based on this, finite element analysis of the active air suspension was performed.

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