A new energy vehicle corner module distributed chassis and control method

Abstract

The present application relates to the technical fields of new energy vehicle distributed chassis, and provides a new energy vehicle angle module distributed chassis and a control method.The chassis comprises a chassis beam, a subframe and a hub angle module, the hub angle module is connected with the chassis beam through the subframe, and the hub angle module integrates a driving assembly, a braking assembly, a suspension assembly and a steering assembly.In the guide mechanism of the suspension assembly, a steering flange output end of a steering driving unit is coaxially arranged with a kingpin axis, and an outer ring connecting flange is connected between a steering knuckle and a steering bearing shell, so that the steering driving unit serves as a guide force transmission component of the suspension assembly to participate in wheel jump guiding and road surface force transmission.The axis of a spring shock absorber assembly is arranged to deviate from a wheel center plane, and an upper swing arm and a lower swing arm are eccentrically hinged relative to the spring shock absorber assembly.The present application can realize independent drive-by-wire collaborative control of wheel driving, braking, steering and suspension.

Description

Technical Field

[0001] This invention relates to the field of distributed chassis technology for new energy vehicles, and in particular to a corner module distributed chassis and control method for new energy vehicles. Background Technology

[0002] As the new energy vehicle industry evolves towards electrification and intelligence, the chassis, as the core carrier of motion control, is undergoing a technological leap from mechanical chassis to intelligent chassis. Early new energy vehicles mostly used the architecture of gasoline vehicles, which limited battery layout and range improvement, and had shortcomings in thermal management and collision safety. While subsequent dedicated electric chassis platforms improved space utilization, their drive, braking, and steering systems still used mechanical or hydraulic connections, resulting in issues such as response delays and insufficient control precision, making it difficult to meet the stringent requirements of advanced autonomous driving.

[0003] The introduction of drive-by-wire technology and distributed drive has opened up new paths for intelligent chassis. Solutions exemplified by skateboard chassis, through drive-by-wire, braking, and steering, have initially achieved decoupling between the chassis and the vehicle body. Distributed drive technology places the motor at the wheel rim or hub, enabling independent control of the driving force of each wheel. However, existing drive-by-wire chassis still have significant bottlenecks: the four major systems—drive, braking, steering, and suspension—mostly adopt a "distributed axle structure" model, with each system relatively independent in space and mechanically interconnected through axles or subframes. This model results in low integration, large space occupation, insufficient lightweighting, and the inability to achieve independent, coordinated, and precise control of the degrees of freedom of individual wheels. Consequently, the vehicle's power, stability, and maneuverability under complex operating conditions are all constrained, failing to fully unleash the performance potential of an intelligent chassis.

[0004] To address the aforementioned issues, several improvements have been proposed in related fields. Chinese patent application CN118528747A discloses a distributed chassis and control method for electric vehicles, employing a fully free-degree-of-freedom mounting platform design to achieve yaw and roll stability control, suitable for preventing sideslip and rollover in large trucks. Chinese patent application CN109501590B discloses a longitudinally distributed wheeled chassis, achieving longitudinal power transmission between the front and rear wheels through an angled transmission box and drive shaft. Chinese patent application CN118372628A discloses a distributed chassis for transport vehicles suitable for strong magnetic field environments, involving the arrangement of systems such as the frame, power battery, drive transmission, hydraulics, electromagnetic shielding, and superstructure platform.

[0005] In addition, the new generation of international chassis architectures generally suffer from defects such as excessive unsprung mass, complex mechanical interfaces, and poor high-speed handling stability. Domestic manufacturers mainly focus on wheel-side or wheel hub drive solutions. Although they achieve independent drive control, the space utilization and integration are still very limited because they have not completely eliminated the axle structure.

[0006] In summary, during this critical window of opportunity for the integrated development of electrification and intelligentization, how to overcome the technical bottlenecks of existing chassis in terms of integration, decoupled control, and collaborative execution, and develop a wheel-end actuator and chassis system that combines high integration, low unsprung mass, excellent handling stability, and plug-and-play functionality, has become a core technical challenge that urgently needs to be tackled in this field. Summary of the Invention

[0007] To address the aforementioned technical problems, this invention provides a corner module distributed chassis and control method for new energy vehicles, aiming to solve the following technical problems existing in the current intelligent drive-by-wire chassis architecture: 1) In existing drive-by-wire chassis systems, the four core functional units of drive, braking, steering and suspension are spatially dispersed, resulting in a large chassis footprint and complex mechanical interfaces; 2) Because the above systems are often coupled through mechanical structures and have a low degree of drive-by-wire, it is difficult to achieve independent and precise control of the driving force, braking force, steering angle and suspension damping / stiffness of a single wheel, which restricts the vehicle's driving performance potential under complex working conditions. 3) Existing highly integrated corner module solutions generally suffer from problems such as excessive unsprung mass, poor high-speed handling stability, and difficulty in balancing space utilization and suspension support stiffness under large-angle steering conditions around the kingpin.

[0008] This invention discloses a corner module distributed chassis for new energy vehicles, comprising: Chassis frame; The subframe is fixedly connected to the chassis beam; The wheel hub angle module is connected to the chassis beam via the subframe; The hub corner module includes: a drive assembly, a braking assembly, a suspension assembly, and a steering assembly; The suspension assembly includes a guide mechanism and a spring-damper assembly. The guide mechanism includes a support arm, an upper control arm, a lower control arm, and a steering knuckle. The support arm is fixedly connected to the subframe. One end of the upper control arm is pivotally connected to the support arm, and the other end is connected to the steering knuckle through a steering load-bearing housing. One end of the lower control arm is pivotally connected to the support arm, and the other end is connected to the steering knuckle. One end of the spring-damper assembly is connected to the support arm, and the other end is connected to the lower control arm. The steering assembly includes a steering drive unit, which has a steering flange output end connected to the upper arm of the steering knuckle and coaxial with the kingpin axis. The steering drive unit also has an outer ring connecting flange connected between the steering knuckle and the steering load-bearing housing. The steering drive unit is hinged to the upper control arm through the steering load-bearing housing, so that the steering drive unit participates in the guidance of wheel bounce and the transmission of road force as a guiding force transmission component of the suspension assembly. The axis of the spring damper assembly is offset from the center plane of the wheel, the hinge position of the upper control arm and the guide mechanism is eccentrically set relative to the spring damper assembly, and the hinge position of the lower control arm and the lower arm of the steering knuckle is eccentrically set relative to the spring damper assembly.

[0009] Optionally, the spring shock absorber assembly includes a telescopic air spring, a hydraulic active shock absorber, and a shock absorber fork arm; The cylindrical air spring, the hydraulic active shock absorber, and the shock absorber fork arm are arranged coaxially in sequence, and the shock absorber fork arm is connected to the lower swing arm; The hydraulic active shock absorber is provided with a recovery chamber valve body and a compression chamber valve body. The recovery chamber valve body is provided with a recovery chamber solenoid valve, and the compression chamber valve body is provided with a compression chamber solenoid valve.

[0010] Optionally, the steering drive unit includes a torque motor, a harmonic reducer, an encoder, and a brake. The output end of the torque motor is connected to the harmonic reducer, the code disk of the encoder is mounted on the rotor shaft of the torque motor, and the brake is used to lock the rotor shaft.

[0011] Optionally, the braking assembly includes a brake electric drive unit, a hub motor brake disc, and a brake caliper; the brake electric drive unit includes a brake actuator motor, a planetary reducer, a parallel shaft input gear, a parallel shaft output gear, a lead screw, a nut, and a brake cylinder. The output end of the brake actuator motor is connected to the planetary reducer, the output end of the planetary reducer is fixedly connected to the parallel shaft input gear, the parallel shaft input gear meshes with the parallel shaft output gear, the parallel shaft output gear is connected to the lead screw, the nut is installed on the lead screw and fixed to the brake cylinder, and the end of the brake cylinder contacts the brake caliper.

[0012] Optionally, the hydraulic active shock absorber can adjust the damping force by adjusting the opening of the solenoid valve of the recovery chamber and the solenoid valve of the compression chamber, or actively inject oil into the recovery chamber or compression chamber through an electro-hydraulic pump to output active force.

[0013] Another aspect of the present invention discloses a control method for the aforementioned corner module distributed chassis of new energy vehicles, comprising: Step S1: Collect chassis status signals, which include at least the output torque of the four wheels, the wheel end angle of the four wheels, the vertical acceleration of the unsprung mass, the vertical acceleration of the sprung mass, the suspension travel, the yaw rate, the longitudinal vehicle speed, the lateral vehicle speed, the longitudinal acceleration, the lateral acceleration, and the vertical displacement of the vehicle body. Step S2: Based on the chassis state signal, obtain the chassis pitch angle prediction sequence and chassis roll angle prediction sequence through the vertical dynamics open-loop prediction module, and input the chassis pitch angle prediction sequence and chassis roll angle prediction sequence into the linear quadratic regulator based on the seven-degree-of-freedom vehicle model to obtain the control force vector of the four-wheel hydraulic actuator. Step S3: Based on the chassis status signal, the optimal additional yaw moment is obtained by the lateral model prediction control module that considers all-wheel steering. Based on the desired longitudinal force and the optimal additional yaw moment, the four-wheel drive torque is optimized and distributed by the quadratic programming algorithm to obtain the four-wheel drive torque vector. Step S4: Based on the rear wheel slip angle, obtain the rear wheel steering angle command through the gain-dispatch PI controller; Step S5: Map the control force vector of the four-wheel hydraulic actuator, the four-wheel drive torque vector, and the rear wheel steering angle command to the hydraulic active damper, wheel hub drive motor or brake electric drive unit, and steering drive unit of each wheel hub angle module, respectively.

[0014] Optionally, in step S2, the vertical dynamics open-loop prediction module takes the current longitudinal vehicle speed, lateral vehicle speed, yaw rate, chassis pitch angle, chassis roll angle, four-wheel wheel end angle and four-wheel output torque as inputs, and recursively calculates the chassis pitch angle prediction sequence and the chassis roll angle prediction sequence based on the pitch-roll dynamics equation in the prediction time domain.

[0015] Optionally, in step S3, the lateral model predictive control module considering all-wheel steering treats both the front wheel angle and the rear wheel angle as measurable external disturbance terms, establishes a linear time-varying model predictive control model including the front and rear wheel angles, uses the center of gravity sideslip angle and yaw rate as state variables, and uses the additional yaw moment as control variable.

[0016] Optionally, in step S3, the objective function of the quadratic programming algorithm is to minimize the tire adhesion utilization rate, and the weight matrix of the objective function is constructed based on the reciprocal of the square of the vertical load of each wheel; the equality constraints of the quadratic programming algorithm include at least the yaw moment balance equation and the total longitudinal force balance equation, and the inequality constraints include at least the tire adhesion elliptic limit and the peak torque boundary of the motor or brake actuator.

[0017] Optionally, in step S4, the gain scheduling PI controller uses the centroid sideslip angle as the controlled variable and the zero centroid sideslip angle as the tracking target. The proportional gain and integral gain of the gain scheduling PI controller are scheduled in real time based on the stability margin of the rear wheel tire sideslip angle.

[0018] Compared with the prior art, the present invention has at least the following beneficial effects: 1. The corner module distributed chassis of the present invention has the advantages of high integration and compact space. By integrating the four major systems of drive, braking, steering and suspension into a single wheel hub corner module, a high degree of integration of the chassis execution system is achieved, which greatly reduces the number of parts, optimizes the chassis layout and improves space utilization.

[0019] 2. The corner module distributed chassis of this invention possesses excellent dynamic performance. It innovatively integrates the steering component into the guiding mechanism design of the suspension component, improving space utilization while maintaining suspension support stiffness. Combined with hydraulic active shock absorbers, damping and output active force can be adjusted in real time, significantly improving chassis comfort and dynamic response.

[0020] 3. The corner module distributed chassis of the present invention has high mobility and flexibility, and the wheel hub corner module has large-angle steer-by-wire capability. The whole vehicle can realize special movement modes such as lateral translation, diagonal driving (crab walking), and turning on the spot, which are impossible for traditional vehicles, greatly improving the passability and parking convenience in narrow spaces.

[0021] 4. The corner module distributed chassis of the present invention has modularity and versatility. The wheel hub corner module is connected to the chassis through the standardized mechanical interface of the bracket arm, realizing "plug and play". This allows the same module to be quickly adapted to different vehicle platforms, greatly shortening the development cycle and reducing manufacturing and maintenance costs.

[0022] 5. The corner module distributed chassis of the present invention can realize precise independent control of the wheel ends. Each wheel is an independent intelligent execution unit directly controlled by electrical signals, which can realize independent drive, braking, steering and suspension of each wheel, providing reliable conditions for improving the vehicle's maneuverability and stability. Attached Figure Description

[0023] 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.

[0024] Figure 1This is a front view of the corner module distributed chassis of the new energy vehicle of the present invention; Figure 2 This is a side view of the corner module distributed chassis of the new energy vehicle of the present invention; Figure 3 This is a top view of a steering wheel of the corner module distributed chassis of the new energy vehicle of the present invention; Figure 4 This is a top view of another steering mechanism in the corner module distributed chassis of the new energy vehicle of the present invention; Figure 5 This is an exploded view of the drive components of the corner module distributed chassis of the new energy vehicle of the present invention; Figure 6 This is a front view and a sectional view of the braking electric drive unit of the corner module distributed chassis of the new energy vehicle of the present invention. Figure 7 These are the main view and sectional view of the steering drive unit of the corner module distributed chassis of the new energy vehicle of the present invention; Figure 8 This is a main view and schematic diagram of the hydraulic active shock absorber of the corner module distributed chassis of the new energy vehicle of the present invention. Figure 9 This is a front view of the suspension components of the corner module distributed chassis of the new energy vehicle of the present invention; Figure 10 This is a front view of the guiding mechanism of the corner module distributed chassis of the new energy vehicle of the present invention; Figure 11 This is a side view of the chassis of the corner module distributed chassis of the new energy vehicle of the present invention; Figure 12 This is a perspective view of the chassis of the corner module distributed chassis of the new energy vehicle of the present invention; Figure 13 This is a top view of the chassis of the corner module distributed chassis of the new energy vehicle of the present invention, showing its lateral driving direction.

[0025] Explanation of reference numerals in the attached figures: 1-Tire, 2-Steering load-bearing housing, 201-Steering load-bearing housing fork arm, 202-Housing cable outlet hole, 3-Coaster arm, 301-Coaster arm main plate, 302-Coaster arm top, 4-Cylinder air spring, 401-Air spring top bolt, 402-Air spring inlet / outlet, 5-Upper swing arm, 501-Inner point of upper swing arm, 502-Outer point of upper swing arm, 6-Steering drive unit, 601-Steering drive unit electrical interface, 602-Steering fixing bolt, 603-Steering flange bolt, 604-Steering flange output end, 605-Steering flange output end bolt hole, 606-Outer ring connecting flange, 611-Torque motor, 612-Harmonic reducer, 613-Brake, 614-Encoder, 615-Rotor shaft 616-Brake resistor, 617-Steering control unit, 6141-Code disk, 6142-Sensing device, 7-Hydraulic active shock absorber, 701-Restoration chamber oil circuit interface, 702-Restoration chamber valve body, 7021-Restoration chamber solenoid valve, 7022-Restoration chamber check valve, 7023-Throttle orifice, 703-Compression chamber oil circuit interface, 704-Compression chamber valve body, 7041-Compression chamber solenoid valve, 7042-Compression chamber check valve, 705-Aggregator, 70 6-Restoration chamber, 707-Compression chamber, 708-Piston rod, 709-Electro-hydraulic pump, 710-Solenoid valve electrical wiring harness interface, 8-Lower control arm, 801-Lower control arm inner point, 802-Shock absorber hinge point, 803-Lower control arm outer point threaded hole, 9-Wheel hub drive motor, 901-Wheel hub motor rotor, 902-Wheel hub motor flange bolt, 903-Wheel hub motor stator, 904-Wheel hub motor brake disc, 905-Brake caliper, 10-Brake electric drive unit, 1 01-Brake actuator motor, 102-Planetary reducer, 103-Parallel shaft input gear, 104-Parallel shaft output gear, 105-Lead screw, 106-Nut, 107-Brake cylinder, 11-Steering ball joint, 12-Shock absorber fork arm, 1201-Axial locking mechanism, 1202-Fork arm fastening bolt, 13-Height sensor, 14-King pin shaft, 15-Wheel rim, 16-Steering knuckle, 1601-Steering knuckle mounting plate, 1602-Steering knuckle upper arm 1603-Steering knuckle lower arm, 1701-Unsprung acceleration sensor, 1702-Sprung acceleration sensor, 18-Spring damper assembly, 19-Guiding mechanism, 20-Subframe, 203-Front subframe, 204-Rear subframe, 21-Chassis beam, 22-Wheel hub module, 23-Water cooling system, 24-Air pump, 25-Air tank, 26-High voltage battery, 27-High voltage power distribution unit, 28-Low voltage lead-acid battery, 29-Inverter. Detailed Implementation

[0026] 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.

[0027] A specific embodiment of the present invention, such as Figure 1 and Figure 2 As shown, a corner module distributed chassis for a new energy vehicle is disclosed, including a subframe 20, a chassis beam 21, a wheel hub corner module 22, a water-cooled heat dissipation device 23, an air pump 24, an air tank 25, a high-voltage battery 26, a high-voltage power distribution unit 27, a low-voltage lead-acid battery 28, a combined inertial navigation sensor, a vehicle controller, and an inverter 29.

[0028] The chassis beam 21 forms the main load-bearing structure of the chassis along the longitudinal direction; the high-voltage battery 26, as an energy source, is centrally integrated between the chassis beams in a flat package form to optimize the vehicle's center of gravity and load distribution; to meet the requirements of the drive-by-wire suspension, the air pump 24 and the air tank 25 are arranged adjacent to each other at the rear of the chassis to adjust the air pressure of each suspension airbag; the inverter 29 is arranged adjacent to the high-voltage battery 26.

[0029] Power distribution and protection area: The high-voltage power distribution unit 27 is located next to the high-voltage battery 26 and is responsible for the safe distribution of high-voltage power; the low-voltage lead-acid battery 28 is located in a position that is convenient for system wake-up and connection to the low-voltage circuit (such as the front compartment or the side of the chassis) to provide independent power supply protection for the vehicle's low-voltage control system.

[0030] The arrangement described in this invention makes full use of the chassis space freed up by the corner modules, achieving efficient coordination of power, energy storage and control components.

[0031] Furthermore, the hub corner module 22 includes a drive assembly, a braking assembly, a steering assembly, and a suspension assembly.

[0032] Furthermore, the steering knuckle 16 of the drive assembly and the suspension assembly is connected by bolts to provide support for the wheel-end drive assembly by the suspension assembly.

[0033] Furthermore, the drive assembly includes a hub drive motor 9, a wheel rim 15, and a tire 1. The wheel rim 15 is positioned between the hub drive motor 9 and the tire 1. The wheel rim 15 is connected to the hub motor flange bolt 902 via bolt holes. The hub motor flange bolt 902 is mounted on the hub motor rotor 901, allowing the driving or braking force of the hub drive motor 9 to be transmitted to the wheel rim 15, thereby driving the tire 1 to rotate or brake and decelerate. The hub motor stator 903 is bolted to the steering knuckle mounting plate 1601, providing a fixed support function and transmitting the impact of the road surface.

[0034] Furthermore, the braking assembly includes a brake electric drive unit 10, a hub motor brake disc 904, and a brake caliper 905.

[0035] Furthermore, the brake electric drive unit 10 is internally equipped with a brake actuator motor 101. At the output end of the brake actuator motor 101, a planetary reducer 102 and a parallel shaft input gear 103 are sequentially mounted from the outside to the inside. The parallel shaft input gear 103 meshes with a parallel shaft output gear 104. The parallel shaft output gear 104 is further connected to a lead screw 105. A nut 106 is installed on the lead screw 105. The nut 106 is locked to the brake cylinder 107 by threads, and the end of the brake cylinder 107 directly contacts the brake caliper 905.

[0036] Furthermore, the rotor of the brake actuator motor 101 rotates under the drive of the stator winding, outputting braking torque to the planetary reducer 102 to achieve deceleration and torque increase. The output end of the planetary reducer 102 is fixedly connected to the parallel shaft input gear 103 by bolts, so that the torque is transmitted to the parallel shaft gear pair, and then the parallel shaft output gear 104 transmits the torque to the lead screw 105, driving the lead screw 105 to rotate.

[0037] Furthermore, when the parallel shaft output gear 104 drives the lead screw 105 to rotate, the nut 106 moves linearly along the axial direction under the action of the balls, and the brake cylinder 107 moves axially accordingly. The brake cylinder 107 pushes the brake caliper 905 to produce displacement, so that the brake caliper 905 presses against the hub motor brake disc 904, and outputs braking force through friction.

[0038] In addition, the braking torque of the hub motor brake disc 904 is transmitted to the hub motor rotor 901 via bolts, and finally to the wheel to achieve deceleration.

[0039] Further, see Figure 7 The steering assembly is a steering drive unit 6; the steering drive unit 6 includes a steering drive unit electrical interface 601, a steering fixing bolt 602, a steering flange bolt 603, a steering flange output end 604, a steering flange output end bolt hole 605, an outer ring connecting flange 606, a torque motor 611, a harmonic reducer 612, a brake 613, an encoder 614, a rotor shaft 615, a braking resistor 616, and a steering control unit 617.

[0040] Furthermore, the steering flange output end 604 is located at the bottom of the steering drive unit 6. Bolt holes 605 are evenly distributed on the steering flange output end 604. The steering flange output end 604 is connected to the upper arm 1602 of the steering knuckle through the steering flange bolts 603. The steering flange output end 604 is coaxially arranged with the kingpin axis 14, so that the driving torque of the steering drive unit 6 can be efficiently transmitted to the steering knuckle 16. The steering ball joint 11 can provide rotational freedom around the kingpin axis, thereby driving the tire 1 to rotate around the kingpin axis 14 or resist the external excitation of the road surface.

[0041] Furthermore, the steering drive unit 6 is equipped with an outer ring connecting flange 606, which is located between the steering knuckle 16 and the steering load-bearing housing 2. The outer ring connecting flange 606 is connected to the steering load-bearing housing 2 by bolts, thus fixing the steering drive unit 6. This makes the steering drive unit 6 part of the suspension assembly, enabling it to guide wheel movement and transmit forces. This design effectively reduces the vertical space occupied, demonstrating the advantages of integrated design of steering and suspension components.

[0042] Furthermore, the torque motor 611 is responsible for converting electrical energy into mechanical energy to output torque. The torque motor 611 transmits power to the harmonic reducer 612, which plays a role in speed reduction and torque increase. In this example, the reduction ratio of the harmonic reducer is 121, and the encoder 614 is a high-precision position detection element.

[0043] Furthermore, the encoder 614 includes a code disk 6141 and a sensing device 6142. The code disk 6141 is directly mounted on the non-output end of the rotor shaft 615 of the torque motor 611 via an interference fit, ensuring absolute synchronous rotation with the rotor shaft 615. The sensing device 6142 is fixedly installed inside the housing of the steering drive unit 6 and is used to detect the position information of the code disk 6141 in real time, thereby forming a closed-loop feedback of the angular position and speed of the rotor shaft 615, and thus accurately controlling the rotation angle of the tire 1. The brake 613 is a normally closed safety brake device. After current is applied to the brake 613, it can generate braking torque to directly lock the rotor shaft 615, realizing static holding and safe braking, which can effectively resist road impact disturbances when the tire 1 maintains any rotation angle.

[0044] Furthermore, the code disk 6141 is rigidly connected to the motor rotor shaft 615 and rotates synchronously. The code disk has a grid pattern, and the sensing device 6142 is fixed to the housing, maintaining a very small gap with the code disk. When the code disk rotates with the shaft, its physical features periodically change the light field around the sensing device, thereby generating a series of pulse signals that reflect position information.

[0045] Furthermore, the tilt angle formed by the kingpin axis 14 and the vertical plane is set to 5°-12°, preferably 8.2°, so that the tire 1 generates a self-aligning torque during rotation, improving the stability of the chassis at high speeds. The tilt angle is determined by the structure of the guide mechanism 19, specifically by the relative spatial positions of the inner and outer mounting points of the upper control arm 5, the inner and outer mounting points of the lower control arm 8, and the upper hinge center and the lower steering ball joint 11 center of the steering knuckle 16.

[0046] Furthermore, the suspension assembly includes a guide mechanism 19 and a spring damper assembly 18. The guide mechanism 19 includes a support arm 3, an upper control arm 5, a lower control arm 8, a steering knuckle 16, a steering load-bearing housing 2, and a steering ball joint 11.

[0047] Furthermore, the steering load-bearing housing 2 includes a steering load-bearing housing fork arm 201 and a housing cable outlet hole 202; the steering knuckle 16 includes an upper steering knuckle arm 1602, a steering knuckle mounting plate 1601 and a lower steering knuckle arm 1603; the steering knuckle mounting plate 1601 is disposed between the upper steering knuckle arm 1602 and the lower steering knuckle arm 1603; the support arm 3 includes a support arm body plate 301 and a support arm top end 302.

[0048] Further, see Figure 10 The upper control arm 5 is U-shaped, including an inner point 501 and an outer point 502. The inner point 501 of the upper control arm is pivotally connected to the main body plate 301 of the support arm through a bushing. The other end of the upper control arm 5 is the outer point 502 of the upper control arm, which is connected to one end of the steering bearing housing fork arm 201 through a bushing. The other end of the steering bearing housing fork arm 201 is connected to the steering drive unit 6 in sequence through bolts. The steering drive unit 6 is connected to the upper arm of the steering knuckle through the output end 604 of the steering flange. The steering knuckle mounting plate is connected to the stator 903 of the hub motor through bolts. The lower arm of the steering knuckle is connected to the lower control arm 8 through the steering ball joint 11.

[0049] Preferably, the pivot point 501 of the upper control arm and the main support plate 301 are connected at 0.7H of the main support plate 301, where H is the total height; in the preferred embodiment of the suspension assembly, in order to achieve a roll center height of about 100mm, the connection position of the inner point 501 of the upper control arm on the main support plate 301 is specially set.

[0050] Specifically, as shown in the attached diagram, the inner point 501 of the upper swing arm is pivotally connected to the support arm main body plate 301 near its top end via a bushing. The vertical distance from the top end 302 of the support arm to its connection point with the inner point 801 of the lower swing arm is defined as H, then the vertical distance between the inner point 501 of the upper swing arm and the inner point 801 of the lower swing arm is 0.7H. This specific height position of the present invention, together with the spatial coordinates of the outer point 502 of the upper swing arm, the inner point 801 of the lower swing arm, and the center of the threaded hole 803 of the outer point of the lower swing arm, constitutes a defined spatial vector system.

[0051] The synergistic effect of this invention is as follows: the aforementioned height position of the inner point 501 of the upper control arm, in conjunction with the specific length (e.g., 0.36H) of the upper control arm 5, determines that the line connecting the inner point 501 and the outer point 502 of the upper control arm forms an installation angle of approximately 7.5°. This installation angle matches the approximately 4.5° installation angle formed by the lower control arm 8. The intersection of these two spatial vectors' virtual extensions in the vehicle's forward-looking plane is precisely defined at a height of approximately 100mm above the ground; this intersection point is the designed roll center.

[0052] The beneficial effect of systematically and collaboratively designing the dimensions and positions of key hard points (upper control arm inner point 501, upper control arm outer point 502, lower control arm inner point 801, and lower control arm outer point threaded hole 803) is that it can optimally achieve the expected roll center height (100mm) within a given layout space constraint. This provides the vehicle with excellent anti-roll capability and enhances handling stability without excessively affecting ride comfort. The specific height of the upper control arm inner point 501 is one of the indispensable key parameters in this collaborative design.

[0053] Furthermore, the lower control arm 8 is Y-shaped, with a lower control arm outer point threaded hole 803 at one end and a lower control arm inner point 801 at the other end; a shock absorber hinge point is provided between the lower control arm outer point threaded hole 803 and the lower control arm inner point 801; the lower control arm outer point threaded hole 803 is connected to the steering ball hinge 11 by a threaded connection, and the lower control arm 8 and the bottom end of the support arm main plate 301 are connected by a bushing at the lower control arm inner point 801.

[0054] Furthermore, based on the arrangement space, the inclination angle between the line connecting the inner point 501 and the outer point 502 of the upper swing arm and the horizontal plane is designed to be 7.5°, and the inclination angle between the line connecting the inner point 801 and the threaded hole 803 of the outer point of the lower swing arm and the horizontal plane is designed to be 4.5°, to match the requirement of a 100mm tilt center height. The above-mentioned components constitute the double-wishbone type guide mechanism 19. It should be noted that this design innovatively incorporates the steering assembly as part of the guide mechanism 19, further improving the compactness of the corner module structure.

[0055] Further, see Figure 9 The spring damper assembly 18 serves as a suspension damping device, comprising a telescopic air spring 4, a hydraulic active damper 7, and a damper fork arm 12, all coaxially arranged from top to bottom. The lower end of the hydraulic active damper 7 is interference-fitted with the damper fork arm 12 and is equipped with an axial locking mechanism 1201 for applying axial clamping force to the axial locking mechanism 1201. The fork arm fastening bolt 1202 is screwed into the threaded hole of the axial locking mechanism 1201, with its end abutting against the outer surface structure of the axial locking mechanism 1201, thereby providing additional axial locking force.

[0056] Furthermore, the tilt angle of the spring-damper assembly 18 is 15.2°. This angle optimizes the leverage ratio, improving the efficiency of vertical force transmission during wheel bounce. This allows the telescopic air spring 4 and the hydraulic active damper 7 to achieve optimal load response within a compact axial travel, thus improving suspension operating efficiency. Simultaneously, it adapts to the geometric arrangement of the wheel-side corner module. This angle effectively prevents motion interference between the steering load-bearing housing 2, the steering drive unit 6, and the spring-damper assembly 18, especially with the recovery chamber valve body 702, during wheel bounce.

[0057] Furthermore, the shock absorber fork arm 12 is connected to the lower control arm 8 via the shock absorber hinge point 802. It should be noted that, viewed from the side of the chassis, components such as the steering knuckle 16, steering drive unit 6, and kingpin axis 14 are arranged on the same plane as the wheel center. (See [reference]). Figure 4 The spring damper assembly 18 is eccentrically positioned on one side of the wheel center. This design, within the limited space at the wheel end, avoids motion interference between components when the drive assembly, braking assembly, and steering knuckle 16 rotate 90° on one side or 40° on the other around the kingpin axis 14, balancing space efficiency with support rigidity. Figure 3 and Figure 4 As shown.

[0058] It is understandable that, such as Figure 3 and Figure 4 As shown, the eccentricity of the spring damper assembly 18 is configured such that the hinge position of the upper control arm 5 to the guide mechanism 19 is eccentric relative to the spring damper assembly 18, and the hinge position of the lower control arm 8 to the steering knuckle lower arm 1603 of the steering knuckle 16 is eccentric relative to the spring damper assembly 18. Based on this eccentric arrangement of the control arms, after the spring damper assembly (18) is installed, its axis naturally deviates from the wheel center plane, thus cleverly avoiding the drive and braking components in the compact space at the wheel end.

[0059] Furthermore, the top of the telescopic air spring 4 is provided with an air inlet and outlet 402. The internal gas volume of the telescopic air spring 4 can be adjusted by the air pump 24, thereby adjusting the chassis height.

[0060] Furthermore, the cylindrical air spring 4 is provided with a top bolt, which is connected to the top end 302 of the support arm.

[0061] Furthermore, a height sensor 13 is installed between the upper swing arm 5 and the support arm 3 to monitor the chassis height and provide feedback information for chassis height adjustment.

[0062] Preferably, the height sensor 13 is a linkage structure. One end of the height sensor 13 is bolted to the support arm 3, and the other end is bolted to the upper control arm 5. The working principle is that the relative positional change between the upper control arm 5 and the support arm 3 can be detected in real time by the height sensor 13. When the vehicle chassis height changes, the upper control arm 5 moves relative to the support arm 3. This movement is sensed by the linkage-type height sensor 13 and converted into an electrical signal, thereby providing accurate feedback to the chassis height adjustment system.

[0063] Further, see Figure 2 The hydraulic active shock absorber 7 is equipped with a recovery chamber oil circuit interface 701, a recovery chamber valve body 702, a compression chamber oil circuit interface 703, a compression chamber valve body 704, and an energy accumulator 705. The hydraulic active shock absorber 7 also has two internal chambers: a recovery chamber 706 and a compression chamber 707. When the piston rod 708 inside the hydraulic active shock absorber 7 moves due to road bumps, it squeezes the oil in the recovery chamber 706 or the compression chamber 707. Guided by the recovery chamber solenoid valve 7021 and the compression chamber solenoid valve 7041, this oil flows through the throttling orifice inside the solenoid valve, such as the throttling orifice 7023 inside the recovery chamber solenoid valve. The control system adjusts the current flowing into the solenoid valve through the solenoid valve electrical harness interface 710, quickly adjusting the opening of the throttling orifice to change the resistance to oil flow, thereby adjusting the damping force.

[0064] Preferably, components such as the recovery chamber oil circuit interface 701, the recovery chamber valve body 702, the compression chamber oil circuit interface 703, the compression chamber valve body 704, and the solenoid valve electrical harness interface 710 are all integrated in a manner perpendicular to the body (outer cylinder) of the hydraulic active shock absorber 7. All interfaces and harnesses are led out radially from the body, avoiding entanglement or interference with peripheral components such as the telescopic air spring 4 and the steering knuckle 16 during the up-and-down reciprocating motion of the hydraulic active shock absorber 7, thus improving reliability.

[0065] Furthermore, an unsprung acceleration sensor 1701 is installed on the steering knuckle 16, and an unsprung acceleration sensor 1702 is installed on the trailing arm 3. These provide acceleration feedback information for the suspension controller to adjust the damping force in real time and improve chassis comfort. During system operation, the sensing components first collect vehicle status signals in real time: the unsprung acceleration sensor 1701 installed on the steering knuckle 16 detects changes in the acceleration of the unsprung mass; the unsprung acceleration sensor 1702 installed on the trailing arm 3 detects changes in the vehicle body's acceleration; and the height sensor 13 installed between the upper control arm 5 and the trailing arm 3 detects changes in suspension travel. These sensors transmit the collected signals to the suspension controller. The suspension controller processes and calculates the received signals, generates control commands, and outputs corresponding voltages to drive the solenoid valves of the hydraulic active shock absorber 7. (See [link to relevant documentation]). Figure 8 and Figure 12 It is driven to act in one or a combination of the following two modes: Real-time damping adjustment mode: In this mode, the suspension controller adjusts the current to the solenoid valve 7021 in the recovery chamber valve body 702 and the solenoid valve 7041 in the compression chamber valve body 704 to change the valve opening, thereby steplessly adjusting the damping force of the oil flowing through them and realizing rapid attenuation of high-frequency excitation from the road surface.

[0066] Active power actuation mode: In this mode, the suspension controller controls the output speed of the electro-hydraulic pump 709. The electro-hydraulic pump 709 draws oil from the accumulator 705 and pressurizes it for output. Simultaneously, the controller coordinates with the solenoid valves 7021 and 7041 to selectively and actively inject high-pressure oil into the recovery chamber 706 or compression chamber 707 of the hydraulic active shock absorber 7, thereby driving the piston rod 708 to output an active thrust or pull force independent of the direction of movement, so as to strongly suppress large changes in the vehicle body posture.

[0067] Furthermore, the hydraulic active damper 7 can also achieve active actuation. After the electro-hydraulic pump 709 is driven to rotate by high-voltage electricity, it draws in oil from the accumulator 705 and pressurizes it. The recovery chamber valve body 702 and the compression chamber valve body 704 actively guide the high-pressure oil from the pump outlet to the recovery chamber 706 or compression chamber 707 of the hydraulic active damper 7 according to the required output force direction, while simultaneously guiding the oil from the other chamber back to the accumulator, thereby rapidly pushing the piston rod 708. By controlling the rotational speed of the electro-hydraulic pump 709 and the opening of the recovery chamber solenoid valve 7021 and the compression chamber solenoid valve 7041, the flow rate and pressure entering the chamber of the hydraulic active damper 7 are actively controlled, thereby generating a precise active power independent of the piston rod 708 speed, used to actively support the vehicle body or suppress movement. In addition, in this example, the electro-hydraulic pump 709 can reach a maximum speed of 7000 rpm and a power of 6.5 kW, which can provide sufficient hydraulic pressure to the suspension system to support the chassis to achieve "dancing" and "jumping" functions.

[0068] Furthermore, the steering drive unit 6, the hydraulic active shock absorber 7, the wheel hub drive motor 9, and the brake electric drive unit 10 are connected to the support arm 3 through the guide mechanism 19 to form a wheel hub corner module 22 with driving, braking, steering, and shock absorption functions.

[0069] Preferably, see Figure 5 and Figure 6 The brake electric drive unit 10 is bolted to the steering knuckle 16, and its specific operation is described in the following section: The brake electric drive unit 10 contains a brake actuator motor 101. At the output end of the brake actuator motor 101, a planetary reducer 102 and a parallel shaft input gear 103 are sequentially mounted from the outside in. The parallel shaft input gear 103 meshes with a parallel shaft output gear 104. The parallel shaft output gear 104 is further connected to a lead screw 105. A nut 106 is mounted on the lead screw 105, and the nut 106 is threadedly locked to the brake cylinder 107. The end of the brake cylinder 107 directly contacts the brake caliper 905.

[0070] Furthermore, the rotor of the brake actuator motor 101 rotates under the drive of the stator winding, outputting braking torque to the planetary reducer 102 to achieve deceleration and torque increase. The output end of the planetary reducer 102 is fixedly connected to the parallel shaft input gear 103 by bolts, so that the torque is transmitted to the parallel shaft gear pair, and then the parallel shaft output gear 104 transmits the torque to the lead screw 105, driving the lead screw 105 to rotate.

[0071] Furthermore, when the parallel shaft output gear 104 drives the lead screw 105 to rotate, the nut 106 moves linearly along the axial direction under the action of the balls, and the brake cylinder 107 moves axially accordingly. The brake cylinder 107 pushes the brake caliper 905 to produce displacement, so that the brake caliper 905 presses against the hub motor brake disc 904, and outputs braking force through friction.

[0072] In addition, the braking torque of the hub motor brake disc 904 is transmitted to the hub motor rotor 901 via bolts, and finally to the wheel to achieve deceleration.

[0073] Furthermore, the wheel hub corner module 22 is bolted to the subframe 20 via the bracket 3, and the subframe 20 is bolted to the chassis beam 21 via high-strength bolts. This allows the driving, braking, steering, shock absorption functions of the wheel hub corner module 22, as well as the road impact load, to be transferred to the chassis beam 21. This single mechanical interface connection method allows the wheel hub corner module 22 to be quickly installed on the chassis and has high versatility, adapting to different chassis.

[0074] Furthermore, for further optimization of system performance and layout, see [link to relevant documentation]. Figure 13 The subframe 20 is designed as a multi-functional integrated platform. To shorten wiring harnesses and improve response speed, key electronic control components such as the braking resistor 616, steering control unit 617, and electro-hydraulic pump 709 are located close to the subframe 20. A water-cooled heat dissipation device 23 is installed on the front subframe 203 to provide precise cooling based on the temperature of the wheel hub drive motor 9. An air pump 24 and an air tank 25 are installed on the rear subframe 204, which, together with the telescopic air spring 4, form a closed air supply circuit to achieve intelligent adjustment of suspension stiffness and chassis height.

[0075] Further, see Figure 11 The chassis's energy and power management unit is located on the chassis beam 21. The high-voltage battery 26 and high-voltage power distribution unit 27 are responsible for centralized high-voltage power supply and management to the wheel hub drive motor 9 and the hydraulic active shock absorber 7. The low-voltage lead-acid battery 28 and inverter 29 together constitute the low-voltage power system. The inverter 29 converts high-voltage DC to low-voltage DC, supplying power to all low-voltage components and controllers of the vehicle while also charging the low-voltage lead-acid battery 28, ensuring efficient energy distribution and safe operation of the system. The highly integrated and modular design of the wheel hub corner modules greatly optimizes the chassis space layout.

[0076] Furthermore, the hub angle module 22 has the function of large-angle steer-by-wire, breaking through the angle limitation of traditional steering systems, enabling the chassis to have high mobility functions such as lateral driving, crab driving, and stationary turning, greatly improving the passability and parking convenience in narrow spaces.

[0077] This invention considers the coordinated design of all mechanisms including drive, braking, steering, and suspension, enabling the hub angle module to achieve a large-angle steer-by-wire capability of 90°. (1) The flat design of the hub drive motor / brake electric drive unit avoids occupying too much lateral space inside the wheel; (2) The installation position of the brake electric drive unit must be such that it can avoid collision with the suspension components when turning at 90°; (3) The steering drive unit must be designed with sufficient power output performance to meet the return torque of the tires when the output torque exceeds 90° of steering. (4) The suspension components should be arranged outside the sweep space of the wheel-end mechanism from -40° to +90°.

[0078] The above embodiments are only used to illustrate the technical solutions of the present invention, and are not intended to limit it. Although the present invention has been described in detail with reference to the foregoing embodiments, those skilled in the art should understand that modifications can still be made to the technical solutions described in the foregoing embodiments, or equivalent substitutions can be made to some or all of the technical features therein. Such modifications or substitutions do not cause the essence of the corresponding technical solutions to deviate from the scope of the technical solutions of the embodiments of the present invention, and they should all be covered within the scope of the claims and specification of the present invention.

[0079] On the other hand, this invention also proposes a chassis control method for the coordinated control of the aforementioned corner module distributed chassis of new energy vehicles, such as... Figures 11-13 The control method adopts a prediction-optimization-execution closed-loop architecture. It uses the torque signals built into the unsprung acceleration sensor 1701, sprung acceleration sensor 1702, height sensor 13, encoder 614, and wheel hub drive motor 9 as control inputs, and the steering drive unit 6, wheel hub drive motor 9, brake electric drive unit 10, hydraulic active shock absorber 7, and telescopic air spring 4 as execution outputs to achieve coordinated optimization of the vehicle's longitudinal, lateral, and vertical dynamics.

[0080] S1: Real-time acquisition and estimation of chassis status signals.

[0081] Specifically, the current output torque of the four wheels (output torque of the front left, front right, rear left, and rear right wheels) is fed back by the controller of the wheel hub drive motor 9. T fl , T fr , T rl , T rr The current wheel end angles of the four wheels (left front, right front, left rear, and right rear wheels) detected by the encoder 614. δ fl , δ fr , δ rl , δ rrThe unsprung mass vertical acceleration of the four wheels is detected in real time by the unsprung acceleration sensor 1701, and the unsprung mass vertical displacement of the four wheels (the unsprung mass vertical displacement of the left front, right front, left rear, and right rear wheels) is obtained by integration. The vertical acceleration of the vehicle body at each wheel end is detected in real time by the sprung acceleration sensor 1702. Suspension travel information is obtained by the height sensor 13; the current yaw rate is obtained by the chassis assembly inertial navigation sensor. Longitudinal speed Lateral speed Longitudinal acceleration a x Lateral acceleration a y Chassis center of gravity height The sideslip angle β is obtained from the sideslip angle estimation formula. The above-mentioned acquired signals are aggregated to the vehicle controller via the vehicle CAN bus, serving as the input source for subsequent prediction and control algorithms.

[0082] Specifically, the expression for the centroid sideslip angle estimation is: .

[0083] S2: For vertical control, a linear quadratic regulator LQR based on fusion model prediction is designed to obtain the lateral control command: the control force vector of the four-wheel hydraulic actuators. It controls the power of the vertical suspension at the ends of the four wheels.

[0084] S21: Construct a fast-computing open-loop prediction module for vertical dynamics, using the current wheel-end angles and output torques of the four wheels as feedforward inputs to predict future tire vertical loads and vehicle attitude, providing key predictive information for collaborative control.

[0085] Specifically, a vertical dynamics open-loop prediction module is constructed, which uses the current longitudinal vehicle speed as the prediction factor. Lateral speed yaw rate Chassis pitch angle Chassis roll angle Four-wheel wheel corner δ fl , δ fr , δ rl , δ rr and the torque output of all four wheels T fl , T fr , T rl , T rrFor input. Within the prediction time domain Np (exemplary, 15 sampling periods, sampling time...), (0.02 s), based on the pitch-roll dynamic equation, the chassis pitch angle prediction sequence is recursively calculated using the Euler method. Chassis roll angle prediction sequence Simultaneously, the vehicle state (longitudinal acceleration, lateral acceleration, vertical displacement of the vehicle body, etc.) at each prediction time is substituted into the dynamic transfer formula of the vertical load to obtain the prediction sequence of the four-wheel vertical load. , It is a 1×4 vector, which contains the vertical loads of the left front, right front, left rear, and right rear wheels in sequence.

[0086] The predicted vertical loads of the four wheels are obtained based on the vehicle's longitudinal and lateral dynamics. This forms the basis for subsequent optimized control and reflects the dynamic transfer of loads. The expression is as follows:

[0087] in, Indicates the first Predicted load value for the left front tire at each step. Indicates the first Predicted load value for the right front tire at each step. Indicates the first Predicted load value for the left rear tire at each step length Indicates the first Predicted load value for the right rear tire at each step length; m For chassis quality, g It is the acceleration due to gravity. , These are the distances from the chassis center of gravity to the front and rear axles, respectively. Indicates the first One step length chassis center of gravity height and These refer to the front track and rear track of the chassis, respectively. Indicates the first Longitudinal acceleration of one step length; Indicates the first Lateral acceleration of one step length.

[0088] Subsequently, based on the longitudinal and lateral velocity components at the center of each of the four wheels, the independent sideslip angles of the four wheels are obtained, expressed as follows:

[0089] in, Indicates the slip angle of the left front tire. Indicates the slip angle of the right front tire. Indicates the left rear tire slip angle. This indicates the right rear tire slip angle.

[0090] Based on the nonlinear correction relationship between vertical load and lateral stiffness, the lateral forces of the four wheels are estimated, expressed as follows:

[0091] in, , , , These represent the lateral forces of the left front, right front, left rear, and right rear tires, respectively. , , , These represent the vertical loads of the left front, right front, left rear, and right rear tires, respectively. , These represent the static loads on the front and rear axles, respectively. , These refer to the lateral stiffness of the front and rear axles of the chassis, respectively.

[0092] Furthermore, pitch and roll dynamic equations are constructed, and pitch moment is determined. With roll moment The expression is:

[0093] in, This is the sum of the longitudinal forces of the four wheels; , For the stiffness of the front axle and rear suspension brackets; , For front and rear suspension damping; I y , I x These are the pitch and roll moments of inertia, respectively. Indicates the chassis pitch rate; Indicates the chassis roll rate; This indicates the chassis pitch angle acceleration; This indicates the chassis roll angle acceleration.

[0094] Furthermore, the chassis pitch angle prediction sequence is obtained through Euler recursion. Chassis roll angle prediction sequence The expression is:

[0095] in, , The first The chassis pitch angle and chassis roll angle are measured in steps. , The first Chassis pitch rate and chassis roll rate in steps.

[0096] Furthermore, the initial values ​​θ(0), φ(0) and... (0) (0) is given by the pitch angle, roll angle and corresponding angular velocity measured by the combined inertial navigation sensor at the current moment.

[0097] This yields the predicted sequences of pitch angle, roll angle, and four-wheel vertical load within the next Np steps. The first value of the predicted sequences of pitch angle, roll angle, and four-wheel vertical load is taken as the predicted value. The predicted sequence of vertical load provides feedforward disturbance compensation for lateral model predictive control and provides real-time boundary conditions for tire adhesion limits for longitudinal force optimization.

[0098] The fast-computing open-loop prediction module for vertical dynamics of this invention is the first to use the independent steering angles and drive torque of the four wheels as feedforward inputs to predict the future vertical loads of the tires and the vehicle attitude in advance, providing key predictive information for coordinated control and overcoming the impact of suspension system response lag on the overall vehicle handling.

[0099] S22: Construct a linear quadratic regulator (LQR) active suspension control strategy based on a 7-DOF vehicle model.

[0100] Chassis pitch angle prediction sequence output by the vertical dynamics open-loop prediction module described in S21 Chassis roll angle prediction sequence As a reference for vehicle attitude, the chassis center of gravity height is detected in real time by the integrated inertial navigation system. The unsprung acceleration sensor 1701 detects the vertical acceleration of the unsprung mass of the four wheels in real time, and the vertical displacement of the unsprung mass of the four wheels is obtained by integration. and .

[0101] A 7-DOF nonlinear vehicle model was established, incorporating vertical, roll, pitch, and four-wheel vertical motions of the vehicle body. The system's generalized coordinate vectors were then used to... Defined as: .

[0102] Furthermore, the equation of motion for the entire vehicle in the vertical direction is:

[0103] in, The vector of the first derivative of the generalized coordinates with respect to time; The vector is the second derivative of the generalized coordinates with respect to time; It is a 7x7 mass matrix, including the sprung mass m s Tilt inertia I xx Pitch inertia Iyy Front axle unsprung mass m uf Rear axle unsprung mass m ur ; This is the stiffness matrix, which includes the front axle suspension stiffness. Rear axle suspension stiffness front axle tire stiffness k tf Rear axle tire stiffness k tr ; It is a 7×7 damping matrix, including the front suspension damping. Rear suspension damping ; To construct the dynamic allocation matrix; This is the control force vector for the four-wheel hydraulic actuator. , The hydraulic actuators control the force for the left front, right front, left rear, and right rear, respectively. Assign a matrix to the perturbation; This represents the road surface disturbance vector.

[0104] Furthermore, select the state vector =[ ] T Construct the state-space equation, the expression of which is:

[0105] in, This represents the transpose of the first derivative vector of the generalized coordinate vector with respect to time. Represents the system matrix; This represents the input matrix.

[0106] Furthermore, the system matrix With input matrix It consists of the following block matrix, expressed as:

[0107] in, Represents a 7×7 zero matrix; Represents a 7×4 zero matrix; This represents a 7×7 identity matrix (with 1s on the main diagonal and 0s on the rest). Furthermore, construct the dynamic allocation matrix. The expression used to describe the coupling effect of dynamics on vehicle body attitude is:

[0108] in, and These are the front track and rear track of the chassis, respectively. This is the longitudinal distance from the front axle to the vehicle's center of gravity. The longitudinal distance from the rear axle to the vehicle's center of gravity.

[0109] Furthermore, to suppress roll and pitch, this invention designs a full-state feedback LQR controller, with the objective function J expressed as:

[0110] in, The semi-positive definite state weight matrix is... ; R q The control weight matrix is ​​positive definite. R q ∈ R 4×4 , R For real numbers, t For time.

[0111] Furthermore, The state weight matrix is ​​a 14x14 positive semi-definite matrix, with dimensions equal to the state vector. Consistent It is a diagonal matrix, and its diagonal elements correspond to the weight coefficients of the vehicle body vertical displacement, chassis roll angle, chassis pitch angle and four-wheel unsprung displacement, respectively. Among them, the weight coefficient corresponding to the vehicle body attitude (roll angle / pitch angle) is greater than the weight coefficient corresponding to the unsprung displacement, so as to prioritize suppressing the change of vehicle body attitude.

[0112] Furthermore, R q It is a 4×4 positive definite control weight matrix, which is a fixed diagonal matrix. Its diagonal elements correspond to the penalty weights of the actuation force of the hydraulic actuators of the left front, right front, left rear, and right rear wheels, respectively. It is used to limit the output amplitude of the actuators and avoid mechanical stroke saturation and excessive impact of the actuators.

[0113] Furthermore, the optimal feedback gain matrix is ​​obtained by solving the algebraic Riccati equation. The expression is:

[0114] in, Let be a 14x14 positive definite symmetric matrix, and be a solution to the algebraic Riccati equation; R is a positive definite control weight matrix q The inverse matrix.

[0115] The optimal control law of the linear quadratic regulator LQR is used as the control force vector of the four-wheel hydraulic actuator. The expression is:

[0116] in, This indicates that the four-wheel suspension feedforward provides power.

[0117] Furthermore, the four-wheel suspension feedforward power... The expression is:

[0118] in, Represents the feedforward assignment matrix; This represents the roll feedforward gain coefficient, used to adjust the sensitivity of roll compensation to the difference in load between the left and right sides of the front axle; This represents the pitch feedforward gain coefficient, used to adjust the sensitivity of pitch compensation to the difference in load between the front and rear axles; , , , These represent the predicted vertical load values ​​for the right front, left front, left rear, and right rear tires in step S21, respectively.

[0119] Furthermore, the vertical load prediction sequence is taken. The first value is used as the predicted vertical load for the right front, left front, left rear, and right rear tires.

[0120] Furthermore, the feedforward assignment matrix The expression is:

[0121] The control force vector of the 4x1 four-wheel hydraulic actuator The control force commands of the four hydraulic actuators, namely the left front, right front, left rear, and right rear, are respectively mapped to the electro-hydraulic pump 709, the solenoid valve 7021 of the recovery chamber, and the solenoid valve 7041 of the compression chamber of the hydraulic active shock absorber 7, driving the piston rod 708 to output active force.

[0122] The linear quadratic regulator LQR active suspension control strategy based on a 7-DOF vehicle model of the present invention suppresses the vertical load transfer caused by the lateral acceleration of the chassis in the corner module.

[0123] S3: Perform lateral model prediction considering all-wheel steering to obtain lateral control commands: optimized four-wheel drive torque vector.

[0124] S31: For the four-wheel independent steering drive unit 6, a lateral model predictive control module considering all-wheel steering is constructed, and the front and rear wheel steering angles are both treated as measurable external disturbance terms, so as to obtain the optimal additional yaw moment in the four-wheel steering driving mode.

[0125] The input to the lateral model predictive control module that considers all-wheel steering is: The parameters include the sideslip angle, yaw rate, wheel-end angles, longitudinal speed, desired sideslip angle (0 in this embodiment), desired yaw rate, and additional yaw torque from the previous sampling step.

[0126] To overcome the limitation of traditional vehicle dynamics models that only consider front-wheel steering, this invention establishes a linear time-varying model predictive control (LTV-MPC) model that includes both front and rear wheel steering angles. Considering the four-wheel independent steering characteristics of the corner module chassis, a linear time-varying model including both front and rear wheel steering angles is established (this step uses a linear two-degree-of-freedom single-track model, merging the left and right front wheels into a front axle and the left and right rear wheels into a rear axle for calculation). The state variable is the sideslip angle. β and yaw rate ω .

[0127] Furthermore, the expression for the linear time-varying model is:

[0128] in, For the overall vehicle weight; It is lateral acceleration; This refers to the longitudinal speed of the vehicle. This refers to the yaw rate; This refers to the lateral force on the front axle. This is the lateral force on the rear axle; This is the common turning angle of the left and right front wheels of the front axle (the turning angles of the left and right front wheels are the same). This is the common turning angle of the left and right rear wheels on the rear axle (the turning angles of the left and right rear wheels on the rear axle are the same). The moment of inertia of the entire vehicle about its vertical axis; Indicates yaw acceleration; This indicates an additional yaw moment.

[0129] Furthermore, , .

[0130] Furthermore, front axle side slip angle and rear axle side slip angle The expression is: .

[0131] Substituting the above expressions for tire lateral force and slip angle into the linear time-varying model, we obtain the continuous state-space expression:

[0132] in, A continuous state vector The first derivative with respect to time, This state vector is the lateral model state vector; Represents the system matrix; Represents the control matrix; E Represents the perturbation matrix. This is the perturbation vector.

[0133] Furthermore, the system matrix Control matrix and perturbation matrix E The derivation is as follows:

[0134] in, The moment of inertia of the entire vehicle about the vertical axis (z-axis); E f This is the column vector of the front wheel steering angle disturbance, i.e., the disturbance matrix. E Column 1; E r This is the column vector of the rear wheel steering angle disturbance, i.e., the disturbance matrix. E The second column.

[0135] Furthermore, for continuous state-space equations Discretization is performed using the forward Euler method; specifically, the system matrix in the equation is discretized. and control matrix Discretization using the forward Euler method yields the discrete system matrix for the linear time-varying model predictive control (LTV-MPC) model that considers the steering angles of the front and rear wheels. and discrete control matrix The expression is:

[0136] Where I is the identity matrix.

[0137] The objective function is defined as minimizing the tracking error and the control increment, expressed as:

[0138] in, For the prediction time domain, i.e., the number of prediction steps; For the first The first step size prediction Each step size outputs a vector; The reference output vector includes the desired centroid sideslip angle and the desired yaw rate; This is the output error weight matrix; To control the time domain, i.e., to control the number of steps; For the first The difference between the additional yaw moment command of each step and the additional yaw moment command of the previous step; R y To control the incremental weight matrix, Indicated by The weighted L2 norm squared is the weight of the weights. Indicates R y It is the squared weighted L2 norm of the weights.

[0139] The output error weight matrix of the present invention Q y =diag(0,10) assigns higher priority to the yaw rate. The objective function J is solved using quadratic programming to obtain the control time domain. N c Control increment sequence within Select the first element of the control increment sequence, i.e., the first element. The difference between the additional yaw moment command for each step and the additional yaw moment command for the previous step. The additional yaw moment of the current step is obtained by superimposing it with the additional yaw moment of the previous sampling step. ,in, Indicates the first The difference between the additional yaw moment command of each step and the additional yaw moment command of the previous step; Indicates the first The difference between the additional yaw moment command of each step and the additional yaw moment command of the previous step.

[0140] S32: Optimize the distribution of drive torque based on the desired vehicle speed and the optimal additional yaw moment; The core of the control method lies in taking the desired longitudinal force output by the upper-level longitudinal controller (which calculates the desired longitudinal force of the whole vehicle through analysis or closed-loop tracking based on the desired vehicle speed given by the driver's accelerator pedal travel or the autonomous driving speed planning module). The optimal additional yaw moment calculated by S31 The four independent hub drive motors 9 or brake electric drive units 10 are allocated to each other through a quadratic programming algorithm.

[0141] Specifically, the quadratic programming algorithm aims to minimize tire adhesion utilization. The weight matrix H of its objective function is constructed based on the reciprocal of the square of the vertical load on each wheel, and its expression is:

[0142] in, The effective radius of the wheel; This represents the road surface adhesion coefficient.

[0143] Furthermore, the optimization objective is: min J =τ T ·H·τ Where τ= [ τfl , τ fr , τ rl , τ rr [ ] represents the four-wheel drive torque vector to be optimized.

[0144] Furthermore, regarding the constraints, the equality constraints must include at least the yaw moment balance equation and the total longitudinal force balance equation, using the small rotation angle approximation cos( δ avg Linearizing the lever arm caused by front wheel steering (≈1), we get:

[0145]

[0146] in, For the desired longitudinal force, This represents the average rotation angle of the front and rear axles. The coefficient matrix represents the constraint of the quadratic programming equation. This represents the right-hand vector of the equality constraint.

[0147] It is understandable that the above As the coefficient matrix of the quadratic programming equality constraints, the As the right-hand vectors of the equality constraints, the two together constitute the constraint equations. ·τ= This ensures that the optimized four-wheel torque vector τ simultaneously satisfies the additional yaw moment. With expected longitudinal force The requirements for overall vehicle dynamics balance.

[0148] Furthermore, the inequality constraints simultaneously consider the tire adhesion elliptic limit and the peak torque boundary of the motor or brake actuator, and are expressed as follows:

[0149]

[0150] in, This represents the lower limit constraint on the torque of the w-th wheel; This represents the upper limit constraint on the torque of the w-th wheel; This represents the vertical load of the w-th wheel, where w takes the values ​​1, 2, 3, and 4, corresponding to the front left, front right, rear left, and rear right, respectively. , These are the negative and positive peak torque limits (e.g., ±800 N·m) for the hub drive motor 9 or the brake electric drive unit 10, respectively.

[0151] By solving the above quadratic programming problem with equality constraints, the optimized four-wheel drive torque vector is obtained and sent to each wheel hub corner module 22 via the high-voltage power distribution unit 27.

[0152] Step S4: Gain scheduling control for rear wheel active steering; To maintain vehicle stability under extreme operating conditions, this invention incorporates a gain-dispatch PI controller based on stability margin in the rear wheel steering channel, which receives real-time observed sideslip angles from the left and right rear wheels. , The left and right rear wheel saturation sideslip angles are obtained by looking up the predicted vertical load values ​​in a table (pre-calibrated based on the sideslip characteristic curve measured on the tire test bench). α srl , α srr This allows us to obtain the rear axle sideslip margin, expressed as:

[0153]

[0154] in, and These represent the stability margin of the left and right rear tire slip angles, respectively. This indicates the stability margin of the rear wheel tire slip angle.

[0155] The controller uses the center of gravity sideslip angle β As the controlled variable, the zero-centimeter sideslip angle is used as the tracking target, and its proportional gain and integral gain are based on the stability margin of the rear wheel tire sideslip angle. Real-time scheduling is performed using the following expression:

[0156] Where, in the formula, The proportional gain is based on the rear tire slip angle stability margin. Real-time scheduling; K p_base Base proportional gain; The integral gain is based on the rear tire slip angle stability margin. Real-time scheduling; Based on the integral gain.

[0157] The rear wheel steering angle command is calculated using the following formula:

[0158] in, This is the error in the centroid sideslip angle. This is a reference value for the centroid sideslip angle. T s The sampling period is This is the original rear wheel steering angle command. 0.087 is the feedforward coefficient; 0.087 is the rear wheel steering angle limit value, in radians. This represents the integral term for the centroid side deflection angle.

[0159] The rear wheel steering angle command, after being limited (e.g., ±5°), is directly output to the rear axle left and right steering drive unit 6, driving the torque motor 611 to rotate the rear wheels around the kingpin axis 14 via the harmonic reducer 612. The common steering angle of the left and right rear wheels is then used to... Receive rear wheel steering angle command.

[0160] Step S5: Multi-actuator collaborative mapping and wire control execution.

[0161] The above longitudinal-lateral-vertical coordinated control commands (longitudinal: four-wheel independent torque command τ=[τ]) will be applied to the longitudinal-lateral-vertical coordinated control commands. fl , τ fr , τ rl , τ rr ] T Lateral: Rear wheel steering angle command; Vertical: Control force vector F of the four-wheel hydraulic actuator. act Finally, the data is mapped to the drive-by-wire actuators of each wheel hub corner module 22 via the vehicle network. Longitudinal control: The independent torque command of the four wheels is converted into three-phase current by the inverter 29 of the hub drive motor 9, which directly drives the wheel rim 15 to output driving force; under braking conditions, the brake actuator motor 101 of the brake electric drive unit 10 pushes the brake cylinder 107 through the planetary reducer 102 and the lead screw nut mechanism, so that the brake caliper 905 presses the hub motor brake disc 904, realizing the coordination of friction braking and energy recovery.

[0162] Lateral control: The front wheel steering angle command and the rear wheel steering angle command are respectively sent to each steering drive unit 6. The torque motor 611 outputs steering torque through the harmonic reducer 612, and the closed-loop angle tracking is achieved through the code disk 6141 of the encoder 614 and the sensing device 6142. The brake 613 locks the rotor shaft 615 after steering is completed to resist the impact of the road surface return torque.

[0163] Vertical control: the chassis pitch angle prediction sequence Chassis roll angle prediction sequence The input is sent to the suspension controller, which, based on the feedback from the unsprung acceleration sensor 1701 and the sprung acceleration sensor 1702, adjusts the opening of the solenoid valve 7021 of the recovery chamber and the solenoid valve 7041 of the hydraulic active shock absorber 7 in real time, and drives the electro-hydraulic pump 709 to actively inject high-pressure oil into the recovery chamber 706 or the compression chamber 707, outputting the main power independent of the movement direction of the piston rod 708, so as to suppress changes in vehicle body posture; at the same time, the air pump 24 adjusts the air pressure of the telescopic air spring 4 based on the feedback from the height sensor 13 to maintain a constant chassis height.

[0164] Through the aforementioned longitudinal-lateral-vertical control method, the corner module distributed chassis incorporates the four independent drive, braking, steering and suspension execution units into a unified model prediction and optimization allocation framework. While making full use of the road surface adhesion potential of each wheel, it achieves precise control of the vehicle's dynamic response and active safety protection under extreme conditions.

Claims

1. A corner module distributed chassis for a new energy vehicle, characterized in that, include: Chassis beam (21); Subframe (20) is fixedly connected to the chassis beam (21); The wheel hub corner module (22) is connected to the chassis beam (21) via the subframe (20); The hub corner module (22) includes: a drive assembly, a braking assembly, a suspension assembly, and a steering assembly; The suspension assembly includes a guide mechanism (19) and a spring damper assembly (18). The guide mechanism (19) includes a support arm (3), an upper control arm (5), a lower control arm (8), and a steering knuckle (16). The support arm (3) is fixedly connected to the subframe (20). One end of the upper control arm (5) is pivotally connected to the support arm (3), and the other end is connected to the steering knuckle (16) through the steering load housing (2). One end of the lower control arm (8) is pivotally connected to the support arm (3), and the other end is connected to the steering knuckle (16). One end of the spring damper assembly (18) is connected to the support arm (3), and the other end is connected to the lower control arm (8). The steering assembly includes a steering drive unit (6), which is provided with a steering flange output end (604). The steering flange output end (604) is connected to the upper arm of the steering knuckle (16), and the steering flange output end (604) is coaxially arranged with the kingpin axis (14). The steering drive unit (6) is also provided with an outer ring connecting flange (606), which is connected between the steering knuckle (16) and the steering bearing housing (2). The steering drive unit (6) is hinged to the upper swing arm (5) through the steering bearing housing (2), so that the steering drive unit (6) participates in the wheel bounce guidance and road force transmission as a guide force transmission component of the suspension assembly. The axis of the spring shock absorber assembly (18) is offset from the center plane of the wheel. The hinge position of the upper swing arm (5) and the guide mechanism (19) is eccentric relative to the spring shock absorber assembly (18). The hinge position of the lower swing arm (8) and the lower arm (1603) of the steering knuckle (16) is eccentric relative to the spring shock absorber assembly (18).

2. The corner module distributed chassis according to claim 1, characterized in that, The spring damper assembly (18) includes a cylindrical air spring (4), a hydraulic active damper (7), and a damper fork arm (12). The cylindrical air spring (4), the hydraulic active shock absorber (7) and the shock absorber fork arm (12) are arranged coaxially in sequence, and the shock absorber fork arm (12) is connected to the lower swing arm (8); The hydraulic active shock absorber (7) is provided with a recovery chamber valve body (702) and a compression chamber valve body (704). The recovery chamber valve body (702) is provided with a recovery chamber solenoid valve (7021), and the compression chamber valve body (704) is provided with a compression chamber solenoid valve (7041).

3. The corner module distributed chassis according to claim 1, characterized in that, The steering drive unit (6) includes a torque motor (611), a harmonic reducer (612), an encoder (614), and a brake (613). The output end of the torque motor (611) is connected to the harmonic reducer (612). The code disk (6141) of the encoder (614) is mounted on the rotor shaft (615) of the torque motor (611). The brake (613) is used to lock the rotor shaft (615).

4. The corner module distributed chassis according to claim 1, characterized in that, The braking assembly includes a brake electric drive unit (10), a hub motor brake disc (904), and a brake caliper (905). The brake electric drive unit (10) includes a brake actuator motor (101), a planetary reducer (102), a parallel shaft input gear (103), a parallel shaft output gear (104), a lead screw (105), a nut (106), and a brake cylinder (107). The output end of the brake actuator motor (101) is connected to the planetary reducer (102), and the output end of the planetary reducer (102) is fixedly connected to the parallel shaft input gear (103). The parallel shaft input gear (103) meshes with the parallel shaft output gear (104), and the parallel shaft output gear (104) is connected to the lead screw (105). The nut (106) is installed on the lead screw (105) and fixed to the brake cylinder (107). The end of the brake cylinder (107) contacts the brake caliper (905).

5. The corner module distributed chassis according to claim 2, characterized in that, The hydraulic active shock absorber (7) adjusts the damping force by adjusting the opening of the solenoid valve (7021) of the recovery chamber and the solenoid valve (7041) of the compression chamber, or actively injects oil into the recovery chamber (706) or the compression chamber (707) through the electro-hydraulic pump (709) to output active force.

6. A control method for the corner module distributed chassis according to any one of claims 1 to 5, characterized in that, include: Step S1: Collect chassis status signals, which include at least the output torque of the four wheels, the wheel end angle of the four wheels, the vertical acceleration of the unsprung mass, the vertical acceleration of the sprung mass, the suspension travel, the yaw rate, the longitudinal vehicle speed, the lateral vehicle speed, the longitudinal acceleration, the lateral acceleration, and the vertical displacement of the vehicle body. Step S2: Based on the chassis state signal, obtain the chassis pitch angle prediction sequence and chassis roll angle prediction sequence through the vertical dynamics open-loop prediction module, and input the chassis pitch angle prediction sequence and chassis roll angle prediction sequence into the linear quadratic regulator based on the seven-degree-of-freedom vehicle model to obtain the control force vector of the four-wheel hydraulic actuator. Step S3: Based on the chassis status signal, the optimal additional yaw moment is obtained by the lateral model prediction control module that considers all-wheel steering. Based on the desired longitudinal force and the optimal additional yaw moment, the four-wheel drive torque is optimized and distributed by the quadratic programming algorithm to obtain the four-wheel drive torque vector. Step S4: Based on the rear wheel slip angle, obtain the rear wheel steering angle command through the gain-dispatch PI controller; Step S5: Map the control force vector of the four-wheel hydraulic actuator, the four-wheel drive torque vector and the rear wheel steering angle command to the hydraulic active damper (7), wheel hub drive motor (9) or brake electric drive unit (10) and steering drive unit (6) of each wheel hub angle module (22).

7. The control method according to claim 6, characterized in that, In step S2, the vertical dynamics open-loop prediction module takes the current longitudinal vehicle speed, lateral vehicle speed, yaw rate, chassis pitch angle, chassis roll angle, four-wheel wheel end angle and four-wheel output torque as inputs, and recursively calculates the chassis pitch angle prediction sequence and the chassis roll angle prediction sequence based on the pitch-roll dynamics equation in the prediction time domain.

8. The control method according to claim 6, characterized in that, In step S3, the lateral model predictive control module considering all-wheel steering treats both the front wheel angle and the rear wheel angle as external disturbance terms, establishes a linear time-varying model predictive control model including the front and rear wheel angles, uses the center of gravity sideslip angle and yaw rate as state variables, and uses the additional yaw moment as control variable.

9. The control method according to claim 6, characterized in that, In step S3, the objective function of the quadratic programming algorithm is to minimize the tire adhesion utilization rate. The weight matrix of the objective function is constructed based on the reciprocal of the square of the vertical load of each wheel. The equality constraints of the quadratic programming algorithm include at least the yaw moment balance equation and the total longitudinal force balance equation, and the inequality constraints include at least the tire adhesion elliptic limit and the peak torque boundary of the motor or brake actuator.

10. The control method according to claim 6, characterized in that, In step S4, the gain scheduling PI controller uses the centroid sideslip angle as the controlled variable and the zero centroid sideslip angle as the tracking target. The proportional gain and integral gain of the gain scheduling PI controller are scheduled in real time based on the stability margin of the rear wheel tire sideslip angle.

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