A method and system for steer-by-wire control optimization based on dual-motor redundant architecture

Abstract

The application discloses a kind of based on double motor redundancy architecture's steer-by-wire control optimization method and system, it is related to steer-by-wire control technical field, including: according to the meshing angle amount of second sun gear in double sun gear different teeth meshing differential mechanism and the meshing angle amount of first sun gear, calculate instantaneous meshing phase;According to the first mechanical power of first motor in steer-by-wire vehicle, the second mechanical power of second motor, calculate internal power difference;According to internal power difference and instantaneous meshing phase, calculate asymmetry coefficient.The application improves the overall efficiency and responsiveness of redundant motor collaborative output.

Description

Technical Field

[0001] This invention relates to the field of steer-by-wire control technology, and in particular to a steer-by-wire control optimization method and system based on a dual-motor redundant architecture. Background Technology

[0002] With the rapid development of intelligent driving and electrification technologies, steer-by-wire chassis systems are increasingly widely used in vehicle control. As a crucial component of steer-by-wire chassis, the steer-by-wire system eliminates the traditional mechanical steering connection structure, instead relying on motors and control units to achieve vehicle steering. In applications such as autonomous driving, unmanned delivery, and special-purpose vehicles, steer-by-wire systems not only require high responsiveness and precision but also redundancy to ensure safe and reliable steering in the event of partial system failures. Against this backdrop, dual-motor redundant drive structures are widely adopted, with differential actuators featuring a dual-sun gear structure being the most prevalent. A dual-sun gear differential actuator typically consists of two drive motors, two sun gears, a co-meshing planetary gear structure, and a rack or pinion connecting the output end, offering advantages such as compactness, redundancy, and high safety. Especially in configurations with unequal numbers of sun gears, a unique co-meshing cycle and load transfer relationship are formed, providing greater design flexibility for the actuator's stability, fault tolerance, and control optimization.

[0003] For the steer-by-wire system with the aforementioned differential structure, existing control strategies generally employ a simple torque distribution method. This method statically distributes torque based solely on the current load capacity or desired output torque of the two motors, neglecting issues such as co-meshing slippage, changes in clamping force, and periodic variations in meshing caused by unequal tooth counts on the sun gear during actual operation. This approach easily leads to uneven force transmission during steering, causing problems such as internal energy circulation, abnormal localized gear stress, and fluctuations in meshing efficiency. Consequently, it affects the stability of the output torque and the service life of the actuator. Summary of the Invention

[0004] The purpose of this invention is to address the shortcomings of existing control strategies that generally employ a simple torque distribution method, which statically distributes torque based solely on the current load capacity or desired output torque of the two motors, neglecting the co-meshing slippage, clamping force changes, and meshing periodic changes caused by the unequal number of teeth on the sun gear during actual operation. This invention proposes a steer-by-wire control optimization method and system based on a dual-motor redundant architecture.

[0005] To address the problems existing in the prior art, the present invention adopts the following technical solution:

[0006] A steer-by-wire control optimization method based on a dual-motor redundant architecture includes:

[0007] S1. Calculate the instantaneous co-meshing phase based on the meshing angle of the second sun gear and the meshing angle of the first sun gear in the dual-sun gear co-meshing differential mechanism.

[0008] S2. Calculate the internal power difference based on the first mechanical power of the first motor and the second mechanical power of the second motor in the steer-by-wire vehicle;

[0009] S3. Calculate the asymmetry coefficient based on the internal power difference and instantaneous co-meshing phase;

[0010] S4. Based on the asymmetry coefficient and the target steering torque of the steer-by-wire vehicle, calculate the basic distribution torque of the first sun gear and the second sun gear respectively;

[0011] S5. Based on the division table of the clamping band section and the basic distribution torque in the double sun gear differential mechanism, determine the final distribution torque of the first sun gear and the second sun gear.

[0012] S6. Optimize the control of the steering actuator of the steer-by-wire vehicle based on the final distributed torque.

[0013] Preferably, the instantaneous co-meshing phase is calculated based on the meshing angle of the second sun gear and the meshing angle of the first sun gear in the dual-sun gear co-meshing differential mechanism, including:

[0014] Positioning the dual-sun differential gear mechanism for steer-by-wire vehicles;

[0015] The meshing angle of the second sun gear is obtained by multiplying the number of teeth of the second sun gear and the rotation angle of the first sun gear in the double sun gear differential mechanism.

[0016] The meshing angle of the first sun gear is obtained by multiplying the number of teeth of the first sun gear and the rotation angle of the second sun gear in the dual-sun gear differential mechanism.

[0017] The phase difference is calculated by measuring the difference between the meshing angle of the second sun gear and the meshing angle of the first sun gear.

[0018] The instantaneous co-meshing phase is obtained by performing a modulo operation on the phase difference.

[0019] Preferably, the internal power difference is calculated based on the first mechanical power of the first motor and the second mechanical power of the second motor in the steer-by-wire vehicle, including:

[0020] Multiply the equivalent torque of the first motor by the angular velocity of the output shaft of the first motor to obtain the first mechanical power output by the first motor to the dual-sun differential gear meshing mechanism;

[0021] Multiply the equivalent torque of the second motor by the angular velocity of the output shaft of the second motor to obtain the second mechanical power output by the second motor to the dual-sun differential gear meshing mechanism;

[0022] The first mechanical power and the second mechanical power are added together to obtain the sum of the input power;

[0023] Obtain the total output torque of the first motor and the second motor after differential synthesis;

[0024] Multiply the total output torque by the angular velocity of the steering rack to obtain the actual mechanical power;

[0025] The difference between the sum of input power and the actual mechanical power is calculated to obtain the internal power difference.

[0026] Preferably, the asymmetry coefficient is calculated based on the internal power difference and the instantaneous co-meshing phase, including:

[0027] The phase space of the dual-sun differential gear meshing mechanism is divided into multiple phase sectors.

[0028] Within each phase sector, the internal power difference is integrated over time along the positive direction of the instantaneous co-meshing phase to obtain the positive rotational energy loss;

[0029] Within each phase sector, the internal power difference is integrated over time along the opposite direction of the instantaneous co-meshing phase to obtain the reverse rotation energy loss.

[0030] The energy difference between forward rotation energy loss and reverse rotation energy loss is calculated to obtain the energy difference.

[0031] The total energy is obtained by adding the energy loss of forward rotation and the energy loss of reverse rotation.

[0032] Dividing the difference energy by the total energy yields the asymmetry coefficient of the phase sector.

[0033] Preferably, based on the asymmetry coefficient and the target steering torque of the steer-by-wire vehicle, the basic distribution torque of the first sun gear and the second sun gear are calculated respectively, including:

[0034] The first participation weight of the first sun gear in the phase sector is calculated based on the asymmetry coefficient.

[0035] Subtracting the first participation weight from 1 yields the second participation weight of the second sun gear;

[0036] Obtain the target steering torque for a vehicle with steer-by-wire control;

[0037] The basic distribution torque of the first sun gear is obtained by multiplying the first participating weight and the steering target torque.

[0038] The basic distribution torque of the second sun gear is obtained by multiplying the second participating weight and the steering target torque.

[0039] Preferably, the formula for calculating the first participating weight is as follows:

[0040] ;

[0041] In the formula, It is the first participation weight of the first sun gear in the i-th phase sector. It is the asymmetry coefficient of the i-th phase sector.

[0042] Preferably, the final distribution torque of the first sun gear and the second sun gear is determined according to the division table of the clamping band section and the basic distribution torque in the dual-sun gear differential mechanism, including:

[0043] Calculate the absolute value of the difference between the number of teeth on the first sun gear and the number of teeth on the second sun gear;

[0044] The absolute value of the difference is defined as the tooth number difference;

[0045] Determine the cycle of co-meshing of different teeth based on the difference in the number of teeth;

[0046] Calculate the periodic position of the instantaneous co-meshing phase within the co-meshing cycle of the opposite-tooth teeth. The formula for calculating the periodic position is as follows:

[0047] ;

[0048] In the formula, It is a periodic position. It is the instantaneous co-meshing phase. It's the difference in the number of teeth. It is a cycle of co-meshing of different teeth;

[0049] Based on the periodic position and the periodicity of the different teeth meshing, a table is established to divide the pressure zone in the dual-sun differential mechanism with different teeth meshing.

[0050] The instantaneous co-meshing phase is classified into different interval categories based on the classification table to obtain the interval category corresponding to the instantaneous co-meshing phase.

[0051] Based on the interval category corresponding to the instantaneous co-meshing phase and the basic distribution torque, the final distribution torque of the first sun gear and the second sun gear is determined.

[0052] Preferably, the final distribution torque of the first and second sun gears is determined based on the interval category corresponding to the instantaneous co-meshing phase and the basic distribution torque, including:

[0053] If the interval category corresponding to the instantaneous co-meshing phase is the first sun gear clamping zone, then the asymmetry coefficient is multiplied by a preset constant to obtain the third participation weight. The result of subtracting the third participation weight from 1 is multiplied by the basic distribution torque of the first sun gear to obtain the final distribution torque of the first sun gear. The final distribution torque of the second sun gear is calculated based on the basic distribution torque of the first sun gear and the second sun gear and the third participation weight.

[0054] If the interval category corresponding to the instantaneous co-meshing phase is the second sun gear clamping zone, then the second motor is unloaded and the unload amount is transferred to the first motor in equal measure to determine the final distribution torque of the first sun gear and the second sun gear.

[0055] If the interval category corresponding to the instantaneous co-meshing phase is the transition zone, then the basic distribution torque of the first sun gear is taken as the final distribution torque of the first sun gear, and the basic distribution torque of the second sun gear is taken as the final distribution torque of the second sun gear.

[0056] Preferably, the formula for calculating the final distributed torque of the second sun gear is as follows:

[0057] ;

[0058] In the formula, It is the final distributed torque of the second sun gear. It is the basic distributed torque of the second sun gear. It is the basic distributed torque of the first sun gear. It is the third participating weight.

[0059] To address the aforementioned problems, this invention also provides a steer-by-wire control optimization system based on a dual-motor redundant architecture, the system comprising:

[0060] The phase calculation module is used to calculate the instantaneous co-meshing phase based on the meshing angle of the second sun gear and the meshing angle of the first sun gear in the dual-sun gear co-meshing differential mechanism.

[0061] The power difference estimation module is used to calculate the internal power difference based on the first mechanical power of the first motor and the second mechanical power of the second motor in the steer-by-wire vehicle.

[0062] The asymmetry coefficient module is used to calculate the asymmetry coefficient based on the internal power difference and instantaneous co-meshing phase.

[0063] The basic distribution module is used to calculate the basic distribution torque of the first sun gear and the second sun gear respectively based on the asymmetry coefficient and the steering target torque of the steer-by-wire vehicle;

[0064] The clamping band decision module is used to determine the final distribution torque of the first sun gear and the second sun gear based on the division table of the clamping band section and the basic distribution torque in the dual sun gear differential meshing mechanism.

[0065] The optimization module is used to optimize the control of the steering actuator of the steer-by-wire vehicle based on the final torque distribution.

[0066] Compared with the prior art, the beneficial effects of the present invention are:

[0067] 1. In this invention, by constructing a co-meshing phase space based on a dual-sun differential gear mechanism, and integrating and comparing the forward rotation energy loss and reverse rotation energy loss in the phase space to form an energy fingerprint table, the energy transmission efficiency under different co-meshing tooth surface states is fully quantified. This can improve the accuracy of torque distribution within the differential structure, avoid energy waste and control lag caused by the unknown output tooth surface state in traditional steer-by-wire systems, and improve the overall efficiency and responsiveness of redundant motor cooperative output.

[0068] 2. In this invention, after the target torque of the differential output is further allocated based on its respective participation weights, a judgment mechanism for the hinge-side off-center load clamping band is introduced. By classifying the clamping band intervals within the co-meshing cycle of the different teeth, and based on the geometric relationship between the periodic position of the instantaneous co-meshing phase and the clamping landing point, it is determined whether the clamping action is concentrated on the tooth surface of the first or second sun gear, thereby selectively adjusting the direction and amplitude of the motor torque distribution. This not only avoids the limitations of torque configuration relying on average distribution or empirical parameters in traditional methods, but also effectively avoids force flow turbulence caused by incorrect clamping direction.

[0069] 3. In this invention, a spatial division and landing point determination strategy of structural periodic quantities is adopted to construct a complete and closed-loop torque optimization process. It does not rely on external sensors to obtain meshing state information, but directly uses the geometric periodic law of the differential structure itself to complete the clamping judgment and participate in weight control. This reduces the system's dependence on high-frequency sensors and algorithm models and improves the system's stability, safety and fault tolerance. Attached Figure Description

[0070] The accompanying drawings, which are included to provide a further understanding of the invention and form part of this application, illustrate exemplary embodiments of the invention and, together with their description, serve to explain the invention and do not constitute an undue limitation thereof. In the drawings:

[0071] Figure 1 This is a flowchart illustrating a steer-by-wire control optimization method based on a dual-motor redundant architecture, as provided in an embodiment of the present invention.

[0072] Figure 2This is a functional block diagram of a steer-by-wire control optimization system based on a dual-motor redundant architecture, provided in an embodiment of the present invention. Detailed Implementation

[0073] The technical solutions of the present invention will be clearly and completely described below with reference to the accompanying drawings of the embodiments of the present invention. Obviously, the described embodiments are only some embodiments of the present invention, and not all embodiments.

[0074] Example: This example provides a steer-by-wire control optimization method based on a dual-motor redundant architecture. See [link to example]. Figure 1 Specifically, including:

[0075] S1. Calculate the instantaneous co-meshing phase based on the meshing angle of the second sun gear and the meshing angle of the first sun gear in the dual-sun gear co-meshing differential mechanism.

[0076] In embodiments of the present invention, the instantaneous co-meshing phase is calculated based on the meshing angle of the second sun gear and the meshing angle of the first sun gear in the dual-sun gear co-meshing differential mechanism, including:

[0077] Positioning the dual-sun differential gear mechanism for steer-by-wire vehicles;

[0078] Specifically, a steer-by-wire vehicle is a vehicle that uses a steer-by-wire system. Its steering system abandons the traditional mechanical connection, collects the driver's steering intention through sensors, processes it through the electronic control unit, and drives the steering actuator with electronic signals to achieve vehicle steering. It has the characteristics of precise steering response, flexible system layout, and easy integration of redundant control strategies, which can effectively improve the safety and intelligence of vehicle steering.

[0079] Specifically, the dual-sun gear differential mechanism is the core component of the steering actuator in a steer-by-wire vehicle. It consists of two sun gears with different numbers of teeth and a planetary gear system that meshes with them. Through the differential meshing motion of the dual sun gears and planetary gears, the output torque of the two motors is synthesized and differentially distributed, thereby providing precise torque transmission and adjustment for the steer-by-wire system and ensuring the stability and redundant control capability of the vehicle steering.

[0080] The meshing angle of the second sun gear is obtained by multiplying the number of teeth of the second sun gear and the rotation angle of the first sun gear in the double sun gear differential mechanism.

[0081] Specifically, the number of teeth on the second sun gear is the total number of teeth on the gear, and the difference in their values ​​forms the structural basis for achieving meshing of different teeth and generating differential effects; the rotation angle of the first sun gear is the angle through which the first sun gear rotates around its axis when it is in operation, reflecting the motion state of the gear; the meshing angle of the second sun gear is a physical quantity obtained by multiplying the number of teeth on the second sun gear by the rotation angle of the first sun gear, and is used to characterize the meshing phase relationship between the second sun gear and the planetary gears.

[0082] The meshing angle of the first sun gear is obtained by multiplying the number of teeth of the first sun gear and the rotation angle of the second sun gear in the dual-sun gear differential mechanism.

[0083] Specifically, the number of teeth on the first sun gear is the total number of teeth on the first sun gear; the rotation angle of the second sun gear is the angle at which the second sun gear rotates around its axis, reflecting the motion state of the gear; the meshing angle of the first sun gear is a physical quantity obtained by multiplying the number of teeth on the first sun gear by the rotation angle of the second sun gear, and is used to characterize the meshing phase relationship between the first sun gear and the planet gears; the meshing angle of the second sun gear is a physical quantity obtained by multiplying the number of teeth on the second sun gear by the rotation angle of the first sun gear, and is used to characterize the meshing phase relationship between the second sun gear and the planet gears.

[0084] The phase difference is calculated by measuring the difference between the meshing angle of the second sun gear and the meshing angle of the first sun gear.

[0085] Specifically, the phase difference is the difference between the meshing angle of the second sun gear and the meshing angle of the first sun gear, and is used to determine the co-meshing phase of the dual sun gears and planetary gears.

[0086] The instantaneous co-meshing phase is obtained by performing a modulo operation on the phase difference.

[0087] Specifically, the steps for performing a modulo operation on the phase difference to obtain the instantaneous co-meshing phase are as follows: First, obtain the difference between the meshing angle of the second sun gear and the meshing angle of the first sun gear as the phase difference value. Then, determine whether the phase difference value is within the range of 0 to 2π. If the phase difference value is less than 0, continuously add 2π to the phase difference value until the result falls within the range of 0 to 2π. If the phase difference value is greater than or equal to 2π, continuously subtract 2π from the phase difference value until the result falls within the range of 0 to 2π. The value obtained after the above adjustment is the instantaneous co-meshing phase, which can accurately reflect the relative meshing positions of the double sun gear and planetary gear in the double sun gear differential mechanism at the current moment.

[0088] Specifically, the instantaneous co-meshing phase is a physical quantity used in the dual-sun gear differential mechanism to characterize the real-time meshing relative position of the first sun gear, the second sun gear, and the planetary gears. It is obtained by equivalent processing of the difference between the meshing angle of the second sun gear and the meshing angle of the first sun gear within the range of 0 to 2π, which can accurately reflect the meshing phase state of the dual sun gears and planetary gears during the operation of the mechanism.

[0089] Specifically, in a dual-sun gear differential meshing mechanism, since the two sun gears have different numbers of teeth, the change in their meshing phase with the planetary gears is directly related to the product of their respective number of teeth and the rotation angle of the other gear. By calculating the product of the number of teeth of the second sun gear and the rotation angle of the first sun gear, and the product of the number of teeth of the first sun gear and the rotation angle of the second sun gear, the meshing angles of the two gears can be obtained. The difference between these two meshing angles reflects the phase difference of the meshing between the dual sun gears and the planetary gears. The modulus operation on this phase difference utilizes the periodicity of the angle to limit it to the range of 0 to 2π, thereby accurately obtaining the relative position of the real-time meshing between the dual sun gears and the planetary gears, i.e., the instantaneous meshing phase. This parameter can accurately characterize the instantaneous state of the meshing between the dual sun gears and the planetary gears during the operation of the mechanism, providing a key phase basis for subsequent phase sector division, torque distribution optimization, etc.

[0090] Specifically, because the two sun gears in the dual-sun differential gear mechanism have different numbers of teeth and mesh with the same set of planetary gears simultaneously, their meshing relationship changes with the phase cycle, thus forming a phase-dependent internal power cycle. This causes the optimal distribution of torque between the two motors under the same target rack force to change with the phase, and when crossing the meshing switching sides, ripple and additional energy consumption caused by slight loosening and re-engagement will occur. Therefore, calculating the instantaneous meshing phase can accurately capture the real-time meshing relative position of the two sun gears and planetary gears during the operation of the mechanism. Based on this, the internal power cycle characteristics under different phases can be analyzed, the optimal torque distribution strategy in each phase interval can be determined, and the phase interval of meshing switching sides can be identified to avoid ripple and additional energy consumption. Ultimately, the consistency, identifiability, and energy transfer efficiency of the torque distribution of the steer-by-wire system under the dual-motor redundant architecture can be optimized.

[0091] S2. Calculate the internal power difference based on the first mechanical power of the first motor and the second mechanical power of the second motor in the steer-by-wire vehicle;

[0092] In an embodiment of the present invention, the internal power difference is calculated based on the first mechanical power of the first motor and the second mechanical power of the second motor in the steer-by-wire vehicle, including:

[0093] Multiply the equivalent torque of the first motor by the angular velocity of the output shaft of the first motor to obtain the first mechanical power output by the first motor to the dual-sun differential gear meshing mechanism;

[0094] Multiply the equivalent torque of the second motor by the angular velocity of the output shaft of the second motor to obtain the second mechanical power output by the second motor to the dual-sun differential gear meshing mechanism;

[0095] Specifically, the first motor and the second motor are two power actuators in the dual-motor redundant architecture steer-by-wire system. They work together to output torque to the dual-sun differential gear mechanism to achieve the synthesis and redundant control of steering torque, providing power support for the steering actuator of the steer-by-wire vehicle, ensuring that the steering function can still be maintained when a single motor fails, and improving energy transmission efficiency by optimizing torque distribution.

[0096] Specifically, the equivalent torque of the first motor is the torque output by the first motor that can be equivalently applied to the double-sun differential gear mechanism, reflecting the torque output capability of the first motor to this mechanism; the output shaft angular velocity of the first motor is the angular velocity of the output shaft, reflecting the speed of its rotation; the first mechanical power is a physical quantity obtained by multiplying the equivalent torque of the first motor by the output shaft angular velocity, characterizing the magnitude of the mechanical power transmitted by the first motor to the double-sun differential gear mechanism. Similarly, the equivalent torque of the second motor is the torque output by the second motor that can be equivalently applied to the double-sun differential gear mechanism, reflecting the torque output capability of the second motor to this mechanism; the output shaft angular velocity of the second motor is the angular velocity of the output shaft, reflecting the speed of its rotation; the second mechanical power is a physical quantity obtained by multiplying the equivalent torque of the second motor by the output shaft angular velocity, characterizing the magnitude of the mechanical power transmitted by the second motor to the double-sun differential gear mechanism.

[0097] The first mechanical power and the second mechanical power are added together to obtain the sum of the input power;

[0098] Obtain the total output torque of the first motor and the second motor after differential synthesis;

[0099] Specifically, the steps to obtain the total output torque are as follows: First, the output torque of the first motor and the output torque of the second motor are collected. Then, based on the tooth number relationship between the first sun gear and the second sun gear in the dual-sun differential gear mechanism and the transmission characteristics of the planetary gear system, the torque of the first motor and the torque of the second motor are differentially synthesized according to the geometric constraints and mechanical transmission laws of the mechanism through the torque synthesis model of the mechanism. Finally, the total output torque of the steering actuator of the steerable vehicle can be obtained. This torque is the final output after the torque of the two motors is coordinated by the differential mechanism, and is used to realize the steering control of the vehicle.

[0100] Specifically, the sum of input power is the result of adding the first mechanical power output by the first motor to the dual-sun differential mechanism and the second mechanical power output by the second motor to the mechanism, representing the total mechanical power input by the two motors to the differential mechanism; the total output torque is the torque obtained after the torques of the first motor and the second motor are differentially synthesized by the dual-sun differential mechanism, which is the final torque output by the mechanism to the steering actuator of the steer-by-wire vehicle, used to drive the vehicle to achieve steering action.

[0101] Multiply the total output torque by the angular velocity of the steering rack to obtain the actual mechanical power;

[0102] The difference between the sum of input power and the actual mechanical power is calculated to obtain the internal power difference.

[0103] Specifically, the steering rack is a key transmission component in the steering actuator of a steer-by-wire vehicle. It works with the steering gear to convert rotational motion into linear motion. One end of the rack receives the total output torque from the dual-sun differential gear mechanism. Through its own linear displacement, it drives the wheels to deflect, thereby achieving the vehicle's steering action.

[0104] Specifically, the angular velocity of the steering rack is the angular velocity of the steering rack's movement, reflecting its speed of motion; the actual mechanical power is a physical quantity obtained by multiplying the total output torque by the steering rack's angular velocity, characterizing the effective mechanical power output by the steering system; the internal power difference is the difference between the sum of the input power and the actual mechanical power, reflecting the ineffective power loss generated within the system due to meshing losses, power cycling, etc.

[0105] Specifically, since the calculation of mechanical power follows the physical law of multiplying torque by angular velocity, the first mechanical power and the second mechanical power are the mechanical power input by the first motor and the second motor to the dual-sun differential gear mechanism, respectively. The sum of the input power is the total power input by the two motors to the mechanism. The total output torque is the output torque after differential synthesis of the torques of the two motors. The actual mechanical power is the effective output power generated by this torque driving the steering rack. According to the physical principle of energy conservation and loss, the difference between the total input power and the effective output power reflects the ineffective power loss generated by relative sliding of gear meshing and power circulation within the system. Therefore, the difference between the sum of the input power and the actual mechanical power is calculated to obtain the internal power difference.

[0106] S3. Calculate the asymmetry coefficient based on the internal power difference and instantaneous co-meshing phase;

[0107] In embodiments of the present invention, the asymmetry coefficient is calculated based on the internal power difference and the instantaneous co-meshing phase, including:

[0108] The phase space of the dual-sun differential gear meshing mechanism is divided into multiple phase sectors.

[0109] Specifically, the phase space is the interval formed by all possible instantaneous co-meshing phases of the dual-sun differential meshing mechanism, reflecting the range of changes in the meshing phase of the mechanism; the phase sector is a number of sub-intervals into which the phase space is equally divided, used for partitioned analysis of the energy loss characteristics of different phase segments.

[0110] Specifically, the steps for dividing the phase space of the dual-sun differential gear meshing mechanism into multiple phase sectors are as follows: First, determine the total angular range of the phase space of the mechanism as 0 to 2π, which is determined by the value characteristics of the instantaneous meshing phase; then, based on the analysis accuracy requirements or control strategy design objectives, set the required number of phase sectors n, where n is a positive integer greater than 1; then calculate the angular interval of each phase sector, that is, by dividing the total angular range 2π by the number of sectors n, the angular width of a single sector is obtained as 2π / n; finally, starting from the starting point 0 of the phase space, divide the boundaries sequentially with an angular width of 2π / n as the interval, forming n continuous and non-overlapping sub-intervals, where the range of the first phase sector is 0 to 2π / n, the range of the second phase sector is 2π / n to 2π×2 / n, and so on, until the range of the nth phase sector is 2π×(n-1) / n to 2π, thus completing the equal division of the phase space. The resulting phase sectors are used to analyze the energy loss and torque distribution characteristics of different phase segments.

[0111] Within each phase sector, the internal power difference is integrated over time along the positive direction of the instantaneous co-meshing phase to obtain the positive rotational energy loss;

[0112] Within each phase sector, the internal power difference is integrated over time along the opposite direction of the instantaneous co-meshing phase to obtain the reverse rotation energy loss.

[0113] Specifically, the steps for obtaining the forward rotation energy loss are as follows: Within each phase sector, first determine the forward direction of the instantaneous co-meshing phase corresponding to that sector, then collect the internal power difference data at each moment within that sector, and then, using time as the integration variable, perform time integration calculation along the forward direction of the instantaneous co-meshing phase on the internal power difference from the start time to the end time of that sector to obtain the forward rotation energy loss within that phase sector; the steps for obtaining the reverse rotation energy loss are as follows: Within each phase sector, first determine the reverse direction of the instantaneous co-meshing phase corresponding to that sector, then collect the internal power difference data at each moment within that sector, and then, using time as the integration variable, perform time integration calculation along the reverse direction of the instantaneous co-meshing phase on the internal power difference from the start time to the end time of that sector to obtain the reverse rotation energy loss within that phase sector.

[0114] Specifically, the forward rotation energy loss is the result of time integration of the internal power difference along the positive direction of the instantaneous co-meshing phase in each phase sector, reflecting the energy loss of the mechanism when rotating forward in that sector; the reverse rotation energy loss is the result of time integration of the internal power difference along the negative direction of the instantaneous co-meshing phase in each phase sector, reflecting the energy loss of the mechanism when rotating in reverse in that sector; the internal power difference is the difference between the total input power of the two motors and the actual effective output power of the steering system, reflecting the ineffective power loss caused by meshing loss, power circulation, etc. within the system.

[0115] The energy difference between forward rotation energy loss and reverse rotation energy loss is calculated to obtain the energy difference.

[0116] The total energy is obtained by adding the energy loss of forward rotation and the energy loss of reverse rotation.

[0117] Dividing the difference energy by the total energy yields the asymmetry coefficient of the phase sector.

[0118] Specifically, the differential energy is the difference between the forward rotation energy loss and the reverse rotation energy loss within the same phase sector, reflecting the degree of difference between the forward and reverse rotation energy losses within that sector; the total energy is the sum of the forward rotation energy loss and the reverse rotation energy loss within the same phase sector, reflecting the scale of the total energy loss caused by the internal power difference within that sector; the asymmetry coefficient is the ratio of the differential energy to the total energy, used to quantitatively characterize the degree of asymmetry between the forward and reverse rotation energy losses within that phase sector, and is a key parameter for subsequent optimization of the dual-motor torque distribution strategy.

[0119] Specifically, since the energy loss of the dual-sun differential gear mechanism varies in different rotation directions, the forward rotation energy loss and the reverse rotation energy loss reflect the energy loss during forward and reverse rotation within the phase sector, respectively. The difference energy is the difference between the two, reflecting the degree of difference in forward and reverse energy loss within the sector. The total energy is the sum of the two, representing the scale of total energy loss within the sector. According to the physical definition of asymmetry, it is the ratio of the difference to the total. Therefore, dividing the difference energy by the total energy yields the asymmetry coefficient, which characterizes the degree of asymmetry in forward and reverse rotation energy loss within the phase sector. This coefficient is a key parameter for optimizing the torque distribution of the dual motors and improving the system's energy transfer efficiency.

[0120] Specifically, since the energy loss and torque distribution of the dual-sun differential gear mechanism are phase-dependent, and the energy loss during forward and reverse rotation is asymmetrical, the asymmetry coefficient of the phase sector can quantify the degree of asymmetry in the forward and reverse rotation energy loss within each phase sector. Based on this, the dual-motor torque distribution strategy can be optimized to ensure the rationality of torque distribution in different phase sectors, reduce internal power loss, improve the energy transmission efficiency and control accuracy of the steer-by-wire system, and ensure the stability and redundancy of vehicle steering.

[0121] S4. Based on the asymmetry coefficient and the target steering torque of the steer-by-wire vehicle, calculate the basic distribution torque of the first sun gear and the second sun gear respectively;

[0122] In an embodiment of the present invention, the basic distribution torque of the first sun gear and the second sun gear is calculated based on the asymmetry coefficient and the target steering torque of the steer-by-wire vehicle, including:

[0123] The first participation weight of the first sun gear in the phase sector is calculated based on the asymmetry coefficient.

[0124] In an embodiment of the present invention, the formula for calculating the first participating weight is as follows:

[0125] ;

[0126] In the formula, It is the first participation weight of the first sun gear in the i-th phase sector. It is the asymmetry coefficient of the i-th phase sector.

[0127] Specifically, the asymmetry coefficient is defined as the ratio of the difference between the internal power generated by the two motors within the phase sector to the sum of the difference and sum of the forward and reverse energies, obtained by integrating the internal power difference along both the forward and reverse directions. This value ranges from -1 to +1, with the sign and magnitude of the value representing the direction and strength of energy transfer towards the first sun gear, respectively; when the positive and negative energies are perfectly symmetrical... When the energy is zero, the contributions of the two sun gears to the output should be equally weighted, with the first participating gear receiving half the weight; when the energy is completely biased towards the first sun gear... When the energy distribution is equal to +1, the optimal allocation should be entirely borne by the first sun gear, with a weight of 1; when the energy is completely biased towards the second sun gear... If the value is negative one, the first sun gear should not participate in this sector, and its weight should be zero. Therefore, the weight must monotonically follow the interval between zero and one. The changes, consistent with the three boundary conditions mentioned above, satisfy the conservation constraint that the sum of the weights of the two gears is one. Furthermore, under the small bias approximation and energy minimization criterion, the simplest and unique mapping satisfying linear response and symmetry is: It equals one half multiplied by one plus Thus, the first participation weight of the first sun gear in the i-th phase sector is obtained.

[0128] Specifically, the first sun gear and the second sun gear are two core gear components in the dual-sun differential gear mechanism. They have different numbers of teeth and are both meshed with the same set of planetary gears. They receive the torque output from the first motor and the second motor respectively, and achieve differential torque synthesis through meshing transmission with the planetary gears. They are the key carriers for transmitting power from the dual motors to the steering actuator. The difference in the number of teeth directly affects the phase change and torque distribution characteristics of the mechanism, and together they support the power output and redundant control function of the steer-by-wire system.

[0129] Subtracting the first participation weight from 1 yields the second participation weight of the second sun gear;

[0130] Obtain the target steering torque for a vehicle with steer-by-wire control;

[0131] Specifically, the first participation weight represents the proportion of the first sun gear's contribution to the total output torque in the current phase sector. The larger the value, the more driving force the first sun gear bears. The second participation weight is one minus the first participation weight, representing the proportion of driving force borne by the second sun gear in the same phase sector. The sum of the two is always one to ensure torque conservation. The steering target torque refers to the expected output torque determined by the steer-by-wire system based on driving commands, vehicle speed, and steering angle requirements. It is the steering torque target that the actuator needs to achieve and is used to indicate the rotation direction and assist strength of the vehicle's front wheels.

[0132] The basic distribution torque of the first sun gear is obtained by multiplying the first participating weight and the steering target torque.

[0133] The basic distribution torque of the second sun gear is obtained by multiplying the second participating weight and the steering target torque.

[0134] Specifically, the basic distribution torque is the initial decomposition result of the target torque by the two sun gears according to the participation weight ratio before considering the tooth difference and the correction of the clamping zone. It reflects the basic driving force level that each motor should bear under the current meshing phase. By multiplying the participation weight and the steering target torque, each drive unit can obtain a preliminary distribution value that matches its current energy efficiency and mechanical state, providing an input basis for the subsequent redistribution and dynamic optimization of the clamping zone.

[0135] Specifically, the target steering torque of the steer-by-wire system represents the total output driving torque required by the vehicle under current operating conditions. It is the resultant force exerted by the two motors on the rack side through the differential mechanism. Since the two sun gears are connected to different motors, the load they bear during meshing should be proportional to their respective participation weights. Based on the principle of torque superposition, the total output torque can be regarded as the weighted sum of the output torques of the two motors, while the participation weights reflect the actual energy contribution and mechanical coupling state of each motor. Multiplying the first participation weight by the target torque yields the basic distributed torque that the first sun gear should bear in this phase segment; similarly, multiplying the second participation weight by the target torque yields the basic distributed torque of the second sun gear. This calculation is equivalent to proportionally decomposing the total output torque to two independent drive sources according to power matching and torque balance conditions, achieving reasonable distribution and coordinated control of energy on the input side of the differential structure, thereby ensuring continuous system output, optimal energy, and stable response.

[0136] Specifically, since the first sun gear and the second sun gear in the dual-sun differential mechanism are connected to the first motor and the second motor respectively, and there is an asymmetry in energy loss during the meshing phase of the mechanism, the basic distribution torque of each gear is calculated separately. The steering target torque can be decomposed proportionally according to their respective participation weights (determined by the phase sector asymmetry coefficient), so that each sun gear bears the driving torque that matches its current energy efficiency and mechanical state. This optimizes the torque distribution of the dual motors, improves the energy transmission efficiency of the differential mechanism, ensures continuous system output, optimal energy, and stable response, and provides an input basis for the subsequent torque redistribution and dynamic optimization in the clamping zone.

[0137] S5. Based on the division table of the clamping band section and the basic distribution torque in the double sun gear differential mechanism, determine the final distribution torque of the first sun gear and the second sun gear.

[0138] In an embodiment of the present invention, the final distribution torque of the first sun gear and the second sun gear is determined according to the division table of the clamping band section and the basic distribution torque in the dual-sun gear differential mechanism, including:

[0139] Calculate the absolute value of the difference between the number of teeth on the first sun gear and the number of teeth on the second sun gear;

[0140] The absolute value of the difference is defined as the tooth number difference;

[0141] Determine the cycle of co-meshing of different teeth based on the difference in the number of teeth;

[0142] Specifically, the unequal tooth meshing period refers to the complete rotational cycle required for the meshing state of the two sun gears in a differential mechanism to return from the initial meshing position to the same relative meshing relationship when their tooth counts are unequal. This period is determined by the difference in the number of teeth between the two gears and represents the repetitive pattern of the meshing phase during relative motion. When the difference in the number of teeth is small, the unequal tooth meshing period is longer, and the meshing state changes slowly; when the difference in the number of teeth increases, the period shortens, and the tooth surface contact distribution is updated more frequently. This period determines the repetitive rhythm of energy transfer, torque distribution, and frictional contact within the system, and is an important geometrical parameter for analyzing phase coupling, pressure band position, and torque ripple in differential mechanisms.

[0143] Calculate the periodic position of the instantaneous co-meshing phase within the co-meshing cycle of the opposite-tooth teeth. The formula for calculating the periodic position is as follows:

[0144] ;

[0145] In the formula, It is a periodic position. It is the instantaneous co-meshing phase. It's the difference in the number of teeth. It is a cycle of co-meshing of different teeth;

[0146] Specifically, the periodic position represents the relative position of the instantaneous meshing phase within the entire non-tooth co-meshing cycle, which is equivalent to converting the continuously changing meshing angle into a periodic interval to identify the cyclical pattern of the meshing state; the tooth number difference is the absolute difference in the number of teeth of the two sun gears, which is the main parameter that determines the length of the non-tooth meshing repetition cycle; the non-tooth co-meshing cycle represents the complete angular range required for the meshing state formed by the unequal number of teeth of the two gears to return from the initial position to the same contact relationship, and its value is equal to 2π divided by the tooth number difference.

[0147] Specifically, this calculation formula uses the instantaneous co-meshing phase of the two sun gears as input variables. Through modulo operation, the continuously increasing phase quantity is reduced to a finite interval within the co-meshing cycle of the opposite gears, thus mapping the relative meshing state within the cycle. In the formula, the instantaneous co-meshing phase represents the actual relative angle between the meshing tooth surfaces of the two gears at the current moment, and is a continuous quantity of the meshing state; the tooth number difference represents the degree of inequality in the number of teeth of the two gears, which determines the angle range required to complete one complete co-meshing cycle of the opposite gears; 2π divided by the tooth number difference is the co-meshing cycle of the opposite gears, representing the total rotation angle that the phase should undergo before the tooth surfaces repeatedly contact; the cycle position is obtained by taking the modulus of the instantaneous phase on this cycle, which is equivalent to locating the current meshing state in the cycle space. This combination of formulas follows the basic laws of angular displacement superposition and periodic functions, enabling the mapping of the continuous rotation process of the mechanical system into a periodic variable. This allows the meshing phase of the differential mechanism to be described in a standardized manner, providing a unified benchmark for subsequent phase sector division and clamping band interval judgment.

[0148] Based on the periodic position and the periodicity of the different teeth meshing, a table is established to divide the pressure zone in the dual-sun differential mechanism with different teeth meshing.

[0149] Specifically, the steps for establishing the division table of the clamping band interval in the double-sun gear differential meshing mechanism based on the periodic position and the inter-gear meshing cycle are as follows: First, calculate the specific value of the inter-gear meshing cycle, i.e., the inter-gear meshing cycle, based on the difference in the number of teeth, and determine the angle range of the entire meshing cycle as 0 to the inter-gear meshing cycle; Second, through bench tests or dynamic simulations, continuously collect the tooth surface normal contact pressure data during the meshing process of the double-sun gear and planetary gear within this inter-gear meshing cycle, with a sampling interval not exceeding 0.01π / ΔN to ensure the continuity and accuracy of the data; Next, analyze the collected contact pressure data, and screen out the angle segments with contact pressure values ​​greater than or equal to a preset pressure threshold (this threshold is 80% of the rated working pressure of the mechanism), mark these angle segments as the clamping band characteristic interval, and mark the remaining angle segments as the transition characteristic interval, wherein the preset pressure threshold is passed through the mechanism. The tooth surface contact fatigue strength is determined by a test under rated load. Then, the angle values ​​corresponding to the above-mentioned clamping band characteristic interval and transition characteristic interval are converted into periodic position ranges. Since the position of the instantaneous co-meshing phase of the periodic position is within the co-meshing cycle of the opposite teeth, its value range is from 0 to the co-meshing cycle of the opposite teeth. Therefore, the angle value of the characteristic interval is directly used as the start and end boundary of the periodic position. Then, the periodic position range of each interval is checked to ensure that the boundaries of adjacent intervals are continuous and do not overlap. If there is overlap, the interval with higher contact pressure value is taken as the effective interval. If there is a gap, the gap part is assigned to the adjacent transition characteristic interval. Finally, the information of all effective intervals is sorted out to form a division table. This table includes the interval number, the start value of the periodic position, the end value of the periodic position, the interval category (clamping band area or transition area), and the corresponding tooth surface contact pressure characteristic description, thus completing the establishment of the clamping band interval division table.

[0150] Specifically, the clamping zone refers to the area of ​​concentrated contact pressure determined by the geometric relationship of the meshing angle of the tooth surfaces during the co-meshing cycle of different teeth. The normal contact force between the tooth surfaces is the largest in this area, and the friction loss and energy transfer are the most significant. The division table is a distribution table of intervals established based on the cycle position and cycle length, used to divide the entire meshing cycle into different segments of the clamping zone and the transition zone. The interval category is used to describe the specific meshing mechanical state corresponding to the instantaneous co-meshing phase. By determining the assignment of the phase in the division table, it can be identified whether the differential mechanism is in the stage of torque enhancement, energy exchange or load release at that moment, providing a basis for subsequent torque redistribution and energy efficiency adjustment.

[0151] The instantaneous co-meshing phase is classified into different interval categories based on the classification table to obtain the interval category corresponding to the instantaneous co-meshing phase.

[0152] Specifically, firstly, the instantaneous co-meshing phase is calculated in real time within the control cycle and converted into a cycle position. The obtained cycle position is then used as an index to input into the compression band interval division table. Secondly, according to the start and end angles and coding order of the intervals recorded in the division table, a sequential search or binary search is performed on the cycle positions. If the cycle position is within the start and end angles of a certain entry, its interval category is temporarily determined as the compression zone, transition zone, or relaxation zone marked by that entry. Subsequently, boundary determination rules, including priority rules and hysteresis threshold rules, are applied to the cycle positions falling at the table boundary, prioritizing the classification of the cycle positions to the side with higher contact force. The classification results of two consecutive control cycles are then further refined. Consistency checks are performed to avoid high-frequency jitter. The classification results of the previous cycle are then compared with the current results. If a transition occurs within the hysteresis width, the previous result is kept unchanged. If the transition exceeds the hysteresis width, the switch is confirmed and a new category is output. The confirmed interval category, along with the timestamp, is written to the circular buffer for subsequent torque redistribution and energy estimation. At the same time, the current classification result is returned as the interval category corresponding to the instantaneous co-meshing phase. Finally, if there are failed entries or interpolation gaps in the table, linear extrapolation is performed to complete the table according to the contact force peak direction of the nearest valid entries, ensuring that a unique and stable interval category determination can be obtained at any cycle position.

[0153] Based on the interval category corresponding to the instantaneous co-meshing phase and the basic distribution torque, the final distribution torque of the first sun gear and the second sun gear is determined.

[0154] In an embodiment of the present invention, the final distribution torque of the first sun gear and the second sun gear is determined based on the interval category corresponding to the instantaneous co-meshing phase and the basic distribution torque, including:

[0155] If the interval category corresponding to the instantaneous co-meshing phase is the first sun gear clamping zone, then the asymmetry coefficient is multiplied by a preset constant to obtain the third participation weight. The result of subtracting the third participation weight from 1 is multiplied by the basic distribution torque of the first sun gear to obtain the final distribution torque of the first sun gear. The final distribution torque of the second sun gear is calculated based on the basic distribution torque of the first sun gear and the second sun gear and the third participation weight.

[0156] In an embodiment of the present invention, the formula for calculating the final distribution torque of the second sun gear is as follows:

[0157] ;

[0158] In the formula, It is the final distributed torque of the second sun gear. It is the basic distributed torque of the second sun gear. It is the basic distributed torque of the first sun gear. It is the third participating weight.

[0159] Specifically, the first sun gear clamping zone is the area where the contact pressure is concentrated when the first sun gear meshes with the planetary gears in a double-sun gear differential mechanism. This zone has a large normal contact force on the tooth surface, resulting in significant energy transfer and frictional losses. The asymmetry coefficient is a parameter that quantifies the degree of asymmetry in energy loss during forward and reverse rotation within the phase sector, reflecting the energy loss difference characteristics of that sector. The preset constant is a fixed coefficient determined based on the mechanical characteristics and energy efficiency experiments of the first sun gear clamping zone, used to adjust the participation weight to optimize torque distribution. The third participation weight is obtained by multiplying the asymmetry coefficient by the preset constant when the first sun gear is clamped, characterizing the adjusted proportion of the first sun gear's participation in torque distribution within this clamping zone. The foundation of the first sun gear... The distribution torque, without considering the clamping zone correction, is the initial decomposition result of the steering target torque of the first sun gear based on the participation weight, reflecting its basic driving force level under the current meshing phase. The final distribution torque of the first sun gear is obtained by subtracting the third participation weight from 1 and multiplying it by its basic distribution torque when the first sun gear is in the clamping zone. It is the actual driving torque borne by the first sun gear after considering the clamping zone characteristics. The final distribution torque of the second sun gear is calculated based on the basic distribution torques of the first and second sun gears and the third participation weight. It is the actual driving torque borne by the second sun gear after considering the clamping zone characteristics of the first sun gear, ensuring the rationality and optimal energy efficiency of the torque distribution of the two sun gears in the clamping zone.

[0160] Specifically, the final torque distribution of the first and second sun gears is calculated by combining their respective base torque distribution and participation weighting coefficients. Firstly, the base torque distribution of the first and second sun gears represents the initial torque distribution value determined by the gear's own geometric characteristics and mechanical state under specific operating conditions. Based on the interval category determined by the instantaneous co-meshing phase, the state of the clamping zone and transition zone will affect their torque distribution participation. If the cycle position is in the clamping zone, the participation weight of the first sun gear is amplified, enhancing its torque distribution; conversely, if the cycle position is in the transition zone, the torque distribution coefficient is adjusted according to the third participation weight, reducing the output of the first sun gear and increasing the output of the second sun gear. Finally, the torque distribution of the second sun gear is based on the base torque distribution of the first sun gear and its corresponding participation weight, combined with coefficient β correction, ensuring torque balance and satisfying system power conservation. This calculation process dynamically adjusts according to the mechanical characteristics of each interval to obtain the final torque distribution of the first and second sun gears, thereby achieving optimized output of the two motors at different operating stages.

[0161] If the interval category corresponding to the instantaneous co-meshing phase is the second sun gear clamping zone, then the second motor is unloaded and the unload amount is transferred to the first motor in equal measure to determine the final distribution torque of the first sun gear and the second sun gear.

[0162] Specifically, the second sun gear clamping zone is the area where the contact pressure is concentrated when the second sun gear meshes with the planetary gears in the double sun gear differential mechanism. This zone has a large normal contact force on the tooth surface, resulting in significant energy transfer and frictional losses. Load reduction involves decreasing the torque load borne by the second motor in the second sun gear clamping zone to match the meshing mechanical characteristics of this area. The load reduction amount is the amount of torque reduced by the second motor. The first motor is a drive unit connected to the first sun gear, used to receive the torque corresponding to the load reduction amount transferred from the second motor. The final distributed torque of the first sun gear is the final driving torque borne by the first motor in the second sun gear clamping zone due to the load reduction received from the second motor. The final distributed torque of the second sun gear is the final driving torque borne by the second motor after load reduction in the second sun gear clamping zone. By reducing the load on the second motor and transferring the load reduction amount equally to the first motor, the torque distribution of the double sun gears can be adapted to the meshing mechanical state of the second sun gear clamping zone, achieving optimal system energy transfer efficiency and torque balance.

[0163] Specifically, when the instantaneous co-meshing phase is determined to be in the clamping zone of the second sun gear through interval category determination, the adjustment coefficient is first calculated. This coefficient is the result of multiplying a preset constant by the absolute value of the asymmetry coefficient of the corresponding phase sector. Next, the final distribution torque of the second sun gear is calculated, which is the difference between the basic distribution torque of the second sun gear and the adjustment coefficient, thereby reducing the load on the second motor. The load reduction is the basic distribution torque of the second sun gear multiplied by the adjustment coefficient. Then, the final distribution torque of the first sun gear is calculated, which is the basic distribution torque of the first sun gear plus the aforementioned load reduction, thereby transferring the load reduction of the second motor to the first motor in equal amounts. Through the above steps, the final distribution torque of the first and second sun gears is finally determined, so that the torque distribution of the dual sun gears is adapted to the meshing mechanical characteristics of the clamping zone of the second sun gear, ensuring optimal energy transfer efficiency and torque balance of the system.

[0164] If the interval category corresponding to the instantaneous co-meshing phase is the transition zone, then the basic distribution torque of the first sun gear is taken as the final distribution torque of the first sun gear, and the basic distribution torque of the second sun gear is taken as the final distribution torque of the second sun gear.

[0165] Specifically, the transition zone is the meshing area between the pressure zone and the engagement zone in a dual-sun gear differential mechanism. In this zone, the tooth surface contact pressure is relatively low, and energy transfer and friction loss are in a transitional state. The basic distribution torque of the first sun gear, without considering pressure zone correction, is the result of the initial decomposition of the steering target torque based on the participation weight, reflecting its basic driving force level in the current meshing phase. The final distribution torque of the first sun gear in the transition zone is directly used as the actual driving torque it bears. Similarly, the basic distribution torque of the second sun gear, without considering pressure zone correction, is the result of the initial decomposition of the steering target torque based on the participation weight, reflecting its basic driving force level in the current meshing phase. The final distribution torque of the second sun gear in the transition zone is directly used as the actual driving torque it bears. Because the meshing mechanical characteristics in the transition zone are relatively stable and no additional adjustment is needed, the basic distribution torque can be directly used as the final distribution torque, ensuring the rationality of torque distribution and system stability.

[0166] Specifically, the interval category corresponding to the instantaneous co-meshing phase reflects the meshing mechanical state of the sun gear and planetary gears in the dual-sun differential gear mechanism, such as the concentration of contact pressure in the clamping zone and the transition of mechanical state in the transition zone. The energy transfer efficiency and friction loss characteristics differ between different intervals. The basic distribution torque is the result of the initial torque decomposition without considering interval characteristics. Determining the final distribution torque of the first and second sun gears based on the interval category corresponding to the instantaneous co-meshing phase and the basic distribution torque allows the torque distribution to adapt to the mechanical and energy characteristics of different meshing intervals, achieving a reasonable torque distribution between the two sun gears, optimizing system energy transfer efficiency, ensuring torque balance and system operational stability, and meeting the requirements of steer-by-wire systems for continuous power output and optimal energy efficiency.

[0167] S6. Optimize the control of the steering actuator of the steer-by-wire vehicle based on the final distributed torque.

[0168] Specifically, the steps for controlling and optimizing the steering actuator of a steer-by-wire vehicle based on the final distributed torque are as follows: First, the final distributed torque of the first and second sun gears is acquired in real time. This torque is the target output torque of the dual motors after correction based on the meshing interval category. Second, the final distributed torque of the first sun gear is converted into the target current command for the first motor, and the final distributed torque of the second sun gear is converted into the target current command for the second motor. The conversion process needs to be combined with the torque-current characteristic curve of the motor (pre-calibrated through bench testing, including the correspondence between torque and current at different speeds) to ensure that the current command and the target torque are accurately matched. Next, the target current command is sent to the motor controller, which drives the first and second motors to output corresponding torques. The torques of the two motors are combined into a driving force for the steering actuator (such as a rack) through a dual-sun differential gear mechanism, pushing the rack to achieve the target steering displacement. Simultaneously, displacement sensors, torque sensors, and speed sensors installed on the steering actuator collect the actual rack displacement, output torque, and actual motor speed in real time, and feed this real-time data back to the control unit. The control unit then compares the actual displacement with the target steering displacement under the current vehicle operating conditions (calculated based on the steering wheel angle signal and vehicle speed signal). The displacement deviation is calculated by comparing the actual output torque with the combined torque corresponding to the final distributed torque. The torque deviation is then calculated. Based on the displacement and torque deviations, the target current commands for the first and second motors are dynamically adjusted using a proportional-integral-derivative (PID) control algorithm. The proportional, integral, and derivative coefficients need to be calibrated according to different interval categories (pressurization zone or transition zone). A higher proportional coefficient is used in the pressurization zone to improve response speed, while a larger integral coefficient is used in the transition zone to reduce steady-state error. When the actual motor speed exceeds a preset safety threshold (based on the motor's rated speed and the mechanism's speed), the system adjusts the torque accordingly. When the mechanical strength is determined or the torque deviation continues to exceed the allowable range (not exceeding 5% of the target torque), the control unit activates the redundancy protection mechanism, reduces the output torque of the dual motors to a safe value, and sends a fault signal to the vehicle controller via the vehicle bus. Finally, the above data acquisition, deviation calculation, command adjustment, and drive control process is continuously executed in a loop until the vehicle completes the current steering action. At the same time, the torque distribution data, motor operating parameters, and actuator response characteristics of each control process are recorded. The PID control parameters and torque-current conversion curve are optimized through offline analysis to further improve the control accuracy, response speed, and energy utilization efficiency of the steering actuator.

[0169] Specifically, this solution uses a dual-motor drive structure where the first motor is connected to the first sun gear and the second motor is connected to the second sun gear, forming a dual-motor redundant architecture. In torque distribution, the two motors can cooperate to bear or transfer torque according to the meshing interval category. For example, when the second motor is unloaded, the first motor can receive an equal amount of torque, ensuring the continuity of power output. At the same time, during the control process, the two motors can dynamically compensate for deviations through PID algorithms. When a motor malfunctions, the system can activate redundancy protection and maintain basic functions. This reflects the redundancy design of the dual motors in power output and fault response, ensuring that the steer-by-wire system can still operate stably when a single motor fails, meeting the reliability requirements of the redundant architecture.

[0170] like Figure 2 The diagram shown is a functional block diagram of a steer-by-wire control optimization system based on a dual-motor redundant architecture, provided by an embodiment of the present invention.

[0171] In this embodiment, the functions of each module / unit are as follows:

[0172] The phase calculation module is used to calculate the instantaneous co-meshing phase based on the meshing angle of the second sun gear and the meshing angle of the first sun gear in the dual-sun gear co-meshing differential mechanism.

[0173] The power difference estimation module is used to calculate the internal power difference based on the first mechanical power of the first motor and the second mechanical power of the second motor in the steer-by-wire vehicle.

[0174] The asymmetry coefficient module is used to calculate the asymmetry coefficient based on the internal power difference and instantaneous co-meshing phase.

[0175] The basic distribution module is used to calculate the basic distribution torque of the first sun gear and the second sun gear respectively based on the asymmetry coefficient and the steering target torque of the steer-by-wire vehicle;

[0176] The clamping band decision module is used to determine the final distribution torque of the first sun gear and the second sun gear based on the division table of the clamping band section and the basic distribution torque in the dual sun gear differential meshing mechanism.

[0177] The optimization module is used to optimize the control of the steering actuator of the steer-by-wire vehicle based on the final torque distribution.

[0178] 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 equivalent substitutions or modifications made by those skilled in the art within the scope of the technology disclosed in the present invention, based on the technical solution and inventive concept of the present invention, should be covered within the scope of protection of the present invention.

Claims

1. A steer-by-wire control optimization method based on a dual-motor redundant architecture, characterized in that, Includes the following steps: S1. Calculate the instantaneous co-meshing phase based on the meshing angle of the second sun gear and the meshing angle of the first sun gear in the dual-sun gear co-meshing differential mechanism. S2. Calculate the internal power difference based on the first mechanical power of the first motor and the second mechanical power of the second motor in the steer-by-wire vehicle; S3. Calculate the asymmetry coefficient based on the internal power difference and instantaneous co-meshing phase; The specific steps for calculating the asymmetry coefficient are as follows: The phase space of the dual-sun differential gear meshing mechanism is divided into multiple phase sectors. Within each phase sector, the internal power difference is integrated over time along the positive direction of the instantaneous co-meshing phase to obtain the positive rotational energy loss; Within each phase sector, the internal power difference is integrated over time along the opposite direction of the instantaneous co-meshing phase to obtain the reverse rotation energy loss. The energy difference between forward rotation energy loss and reverse rotation energy loss is calculated to obtain the energy difference. The total energy is obtained by adding the energy loss of forward rotation and the energy loss of reverse rotation. Dividing the difference energy by the total energy yields the asymmetry coefficient of the phase sector. S4. Based on the asymmetry coefficient and the target steering torque of the steer-by-wire vehicle, calculate the basic distribution torque of the first sun gear and the second sun gear respectively; S5. Based on the division table of the clamping band section and the basic distribution torque in the double sun gear differential mechanism, determine the final distribution torque of the first sun gear and the second sun gear. The specific steps for determining the final torque distribution of the first and second sun gears are as follows: Calculate the absolute value of the difference between the number of teeth on the first sun gear and the number of teeth on the second sun gear; The absolute value of the difference is defined as the tooth number difference; Determine the cycle of co-meshing of different teeth based on the difference in the number of teeth; Calculate the periodic position of the instantaneous co-meshing phase within the co-meshing cycle of the opposite-tooth teeth. The formula for calculating the periodic position is as follows: ； In the formula, It is a periodic position. It is the instantaneous co-meshing phase. It's the difference in the number of teeth. It is a cycle of co-meshing of different teeth; Based on the periodic position and the periodicity of the different teeth meshing, a table is established to divide the pressure zone in the dual-sun differential mechanism with different teeth meshing. The instantaneous co-meshing phase is classified into different interval categories based on the classification table to obtain the interval category corresponding to the instantaneous co-meshing phase. Based on the interval category corresponding to the instantaneous co-meshing phase and the basic distribution torque, determine the final distribution torque of the first sun gear and the second sun gear; S6. Optimize the control of the steering actuator of the steer-by-wire vehicle based on the final distributed torque.

2. The steer-by-wire control optimization method based on a dual-motor redundant architecture according to claim 1, characterized in that, Based on the meshing angles of the second sun gear and the first sun gear in the dual-sun gear differential mechanism, the instantaneous co-meshing phase is calculated, including: Positioning the dual-sun differential gear mechanism for steer-by-wire vehicles; The meshing angle of the second sun gear is obtained by multiplying the number of teeth of the second sun gear and the rotation angle of the first sun gear in the double sun gear differential mechanism. The meshing angle of the first sun gear is obtained by multiplying the number of teeth of the first sun gear and the rotation angle of the second sun gear in the dual-sun gear differential mechanism. The phase difference is calculated by measuring the difference between the meshing angle of the second sun gear and the meshing angle of the first sun gear. The instantaneous co-meshing phase is obtained by performing a modulo operation on the phase difference.

3. The steer-by-wire control optimization method based on a dual-motor redundant architecture according to claim 1, characterized in that, Based on the first mechanical power of the first motor and the second mechanical power of the second motor in the steer-by-wire vehicle, the internal power difference is calculated, including: Multiply the equivalent torque of the first motor by the angular velocity of the output shaft of the first motor to obtain the first mechanical power output by the first motor to the dual-sun differential gear meshing mechanism; Multiply the equivalent torque of the second motor by the angular velocity of the output shaft of the second motor to obtain the second mechanical power output by the second motor to the dual-sun differential gear meshing mechanism; The first mechanical power and the second mechanical power are added together to obtain the sum of the input power; Obtain the total output torque of the first motor and the second motor after differential synthesis; Multiply the total output torque by the angular velocity of the steering rack to obtain the actual mechanical power; The difference between the sum of input power and the actual mechanical power is calculated to obtain the internal power difference.

4. The steer-by-wire control optimization method based on a dual-motor redundant architecture according to claim 2, characterized in that, Based on the asymmetry coefficient and the target steering torque of the steer-by-wire vehicle, the basic distribution torques of the first and second sun gears are calculated, including: The first participation weight of the first sun gear in the phase sector is calculated based on the asymmetry coefficient. Subtracting the first participation weight from 1 yields the second participation weight of the second sun gear; Obtain the target steering torque for a vehicle with steer-by-wire control; The basic distribution torque of the first sun gear is obtained by multiplying the first participating weight and the steering target torque. The basic distribution torque of the second sun gear is obtained by multiplying the second participating weight and the steering target torque.

5. The steer-by-wire control optimization method based on a dual-motor redundant architecture according to claim 4, characterized in that, The formula for calculating the first participating weight is as follows: ； In the formula, It is the first participation weight of the first sun gear in the i-th phase sector. It is the asymmetry coefficient of the i-th phase sector.

6. The steer-by-wire control optimization method based on a dual-motor redundant architecture according to claim 1, characterized in that, Based on the interval category corresponding to the instantaneous co-meshing phase and the basic distribution torque, the final distribution torque of the first and second sun gears is determined, including: If the interval category corresponding to the instantaneous co-meshing phase is the first sun gear clamping zone, then the asymmetry coefficient is multiplied by a preset constant to obtain the third participation weight. The result of subtracting the third participation weight from 1 is multiplied by the basic distribution torque of the first sun gear to obtain the final distribution torque of the first sun gear. The final distribution torque of the second sun gear is calculated based on the basic distribution torque of the first sun gear and the second sun gear and the third participation weight. If the interval category corresponding to the instantaneous co-meshing phase is the second sun gear clamping zone, then the second motor is unloaded and the unload amount is transferred to the first motor in equal measure to determine the final distribution torque of the first sun gear and the second sun gear. If the interval category corresponding to the instantaneous co-meshing phase is the transition zone, then the basic distribution torque of the first sun gear is taken as the final distribution torque of the first sun gear, and the basic distribution torque of the second sun gear is taken as the final distribution torque of the second sun gear.

7. The steer-by-wire control optimization method based on a dual-motor redundant architecture according to claim 6, characterized in that, The formula for calculating the final distribution torque of the second sun gear is as follows: ； In the formula, It is the final distributed torque of the second sun gear. It is the basic distributed torque of the second sun gear. It is the basic distributed torque of the first sun gear. It is the third participating weight.

8. A steer-by-wire control optimization system based on a dual-motor redundant architecture, characterized in that, The system includes: The phase calculation module is used to calculate the instantaneous co-meshing phase based on the meshing angle of the second sun gear and the meshing angle of the first sun gear in the dual-sun gear co-meshing differential mechanism. The power difference estimation module is used to calculate the internal power difference based on the first mechanical power of the first motor and the second mechanical power of the second motor in the steer-by-wire vehicle. The asymmetry coefficient module is used to calculate the asymmetry coefficient based on the internal power difference and instantaneous co-meshing phase. The specific steps for calculating the asymmetry coefficient are as follows: The phase space of the dual-sun differential gear meshing mechanism is divided into multiple phase sectors. Within each phase sector, the internal power difference is integrated over time along the positive direction of the instantaneous co-meshing phase to obtain the positive rotational energy loss; Within each phase sector, the internal power difference is integrated over time along the opposite direction of the instantaneous co-meshing phase to obtain the reverse rotation energy loss. The energy difference between forward rotation energy loss and reverse rotation energy loss is calculated to obtain the energy difference. The total energy is obtained by adding the energy loss of forward rotation and the energy loss of reverse rotation. Dividing the difference energy by the total energy yields the asymmetry coefficient of the phase sector. The basic distribution module is used to calculate the basic distribution torque of the first sun gear and the second sun gear respectively based on the asymmetry coefficient and the steering target torque of the steer-by-wire vehicle; The clamping band decision module is used to determine the final distribution torque of the first sun gear and the second sun gear based on the division table of the clamping band section and the basic distribution torque in the dual sun gear differential meshing mechanism. The specific steps for determining the final torque distribution of the first and second sun gears are as follows: Calculate the absolute value of the difference between the number of teeth on the first sun gear and the number of teeth on the second sun gear; The absolute value of the difference is defined as the tooth number difference; Determine the cycle of co-meshing of different teeth based on the difference in the number of teeth; Calculate the periodic position of the instantaneous co-meshing phase within the co-meshing cycle of the opposite-tooth teeth. The formula for calculating the periodic position is as follows: ； In the formula, It is a periodic position. It is the instantaneous co-meshing phase. It's the difference in the number of teeth. It is a cycle of co-meshing of different teeth; Based on the periodic position and the periodicity of the different teeth meshing, a table is established to divide the pressure zone in the dual-sun differential mechanism with different teeth meshing. The instantaneous co-meshing phase is classified into different interval categories based on the classification table to obtain the interval category corresponding to the instantaneous co-meshing phase. Based on the interval category corresponding to the instantaneous co-meshing phase and the basic distribution torque, determine the final distribution torque of the first sun gear and the second sun gear; The optimization module is used to optimize the control of the steering actuator of the steer-by-wire vehicle based on the final torque distribution.

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