Port heavy load IGV hydraulic braking and steering distributed collaborative control method

Abstract

The application discloses a port heavy-load IGV hydraulic braking and steering distributed cooperative control method and belongs to the technical field of vehicle subsystem joint control. The cooperative control method is based on a port heavy-load IGV hydraulic braking and steering control system, which comprises the following target functions: a slope function algorithm is used for speed planning, the movement of each axle is divided into three stages of acceleration, uniform speed and deceleration, the maximum movement time is taken as the synchronization time by assuming that the acceleration and deceleration are of the same size: the maximum speed of each axle is adjusted: the dynamic equation of each axle is: a PID controller or a feedforward control strategy is used to design i the control input of the axle. The application compensates the inertia and damping characteristics of the system through the feedforward link, simultaneously corrects the tracking error in real time by using the PID feedback link, accurately drives the valve opening of the steering proportional valve, thereby adjusts and controls the hydraulic oil flow and direction flowing to the steering oil cylinder, so that the expected steering dynamic characteristics are achieved.

Description

Technical Field

[0001] This invention relates to the field of vehicle subsystem joint control technology, and more specifically to a distributed cooperative control method for hydraulic braking and steering of heavy-duty IGVs in ports. Background Technology

[0002] Intelligent Guided Vehicles (IGVs) are the core equipment for horizontal transport in automated terminals, responsible for the transfer of containers between the quay and the yard. In particular, heavy-duty IGVs with a carrying capacity exceeding 40 tons directly determine the throughput capacity and operating costs of the entire terminal in terms of their operational efficiency, safety, and reliability.

[0003] Chinese invention patent application CN 119261851A discloses a steerable hydraulic steering braking system and method. The steerable hydraulic steering braking method includes: acquiring control commands from a vehicle controller; the control commands include: a target steering angle, a target angular velocity, and a braking percentage command. Based on the control commands, the steering valve assembly is controlled to adjust the angular velocity and deflect the target steering angle to perform different steering modes. A steering angle difference is obtained based on the actual steering angle and the target steering angle, and an angular velocity difference is obtained based on the actual angular velocity and the target angular velocity; the actual steering angle and actual angular velocity are obtained based on a kingpin angle sensor. A PID control method is used to control the steering valve assembly to adjust the steering angle and angular velocity based on the steering angle difference and the angular velocity difference.

[0004] However, due to their multi-axle steering layout, four-axle IGVs exhibit more complex dynamic characteristics than traditional vehicles. Especially under heavy-load conditions, the inertial coupling effect between the axles is significant. Coupled with the inherent nonlinearity and time-varying parameters of the hydraulic system, the aforementioned PID control algorithms struggle to coordinate the synchronous response of the four steering axes. This control delay and insufficient precision directly lead to deviations in the tracking of steering angles for each axle, causing the vehicle trajectory to deviate from the predetermined path. This is particularly pronounced during high-speed steering or emergency obstacle avoidance, severely limiting the efficiency of four-axle IGVs in modern port operations and failing to meet the requirements for high-precision positioning and smooth operation. Therefore, this invention provides a distributed cooperative control method for hydraulic braking and steering in port heavy-duty IGVs. Summary of the Invention

[0005] This invention provides a distributed cooperative control method for hydraulic braking and steering of heavy-duty IGVs in ports, aiming to solve the problem that existing PID control algorithms are difficult to coordinate the synchronous response of the four steering axes of heavy-duty IGVs in ports, which restricts the efficiency of four-axis IGVs in modern port operations and fails to meet the requirements of high-precision positioning and stable operation.

[0006] The present invention adopts the following technical solution:

[0007] A distributed cooperative control method for hydraulic braking and steering of a port heavy-duty IGV is disclosed. Based on a port heavy-duty IGV hydraulic braking and steering control system, the method includes a hydraulic station, a steering system, a braking system, and a controller. The port heavy-duty IGV is a four-axle IGV. Angle sensors on the kingpins of the steering knuckles of the axles detect and provide feedback on the steering angle of each axle in real time. The controller compares the feedback steering angle with the target steering angle in real time, calculates the deviation using a PID control algorithm, and dynamically adjusts the control signal to the steering proportional valve until the actual angle precisely converges to the target steering angle, thereby achieving precise servo control of the steering process. Specifically, the method includes the following processes:

[0008] I. Define the following objective function: ,in, i For each axis, Maximum speed for each axis, t For the motion time of all axes, yes i Axis control input, T This is the time required for all axes to reach the target steering angle;

[0009] II. Velocity planning using the ramp function algorithm: (1) Divide the motion of each axis into three stages: acceleration, constant speed, and deceleration. Assume that the magnitudes of acceleration and deceleration are the same, and the angle of each axis is . The initial angle is The acceleration of each axis is (2) Calculate the total motion time for each axis. (3) Take the maximum motion time as the synchronization time: (4) Adjust the maximum speed of each axis to (5) The dynamic equations for each axis are: ;in, yes i Moment of inertia of the shaft yes i The damping coefficient of the shaft, yes i The actual angular acceleration of the axis, yes i The actual angular velocity of the shaft;

[0010] III. Design using a PID controller or feedforward control strategy. i Axis control input : ,in, For the desired angle position, For the desired angular velocity, For the desired angular acceleration, For proportional gain, This is the differential gain.

[0011] Specifically, the equation of motion in step (2) of the speed planning process is as follows: ① Acceleration phase : ,in, ② Uniform speed stage : ,in, ③ Deceleration phase : .

[0012] The total motion time for each axis in step two above Represented as: Maximum speed is : , Let be the acceleration time along the i-axis.

[0013] The above-mentioned port heavy-duty IGV hydraulic brake-steering control system realizes the following steering modes, including: dual front axle steering mode suitable for high-speed cruising, dual rear axle steering mode suitable for operation in narrow spaces, and all-wheel steering mode with a very small turning radius.

[0014] The interior angles of the first to fourth axles of the heavy-duty IGV in the port are set as follows: , , , The formulas for calculating the interior angles of each mode are as follows: Dual front axle steering mode: Dual rear axle steering mode: Four-axle steering mode: Diagonal mode: In the formulas for calculating the interior angle of each steering mode above, L1 is the interior angle of the first axis given by the system; L2 is the wheelbase between the first and second axes; L3 is the wheelbase between the second and third axes; L4 is the wheelbase between the third and fourth axes.

[0015] To ensure the most effective control of braking force, the objective function is designed as follows: ;in, For the first Axis in t Braking force at all times for t Total braking force requirement at any moment For the first i Maximum braking force limit of the shaft, For the first i Limitation on the rate of change of braking force on the shaft. , These are the weighting coefficients. t f This represents the total braking time.

[0016] As can be seen from the above description of the present invention, compared with the prior art, the present invention has the following advantages:

[0017] This invention's controller compares the feedback steering angle with the target steering angle in real time, calculates the deviation using a PID control algorithm, and dynamically adjusts the control signal for the proportional valve until the actual angle precisely converges to the target steering angle, thus achieving precise servo control of the steering process. A composite control algorithm using feedforward and PID feedback generates the corresponding control signal, precisely driving the valve opening of the proportional steering valve, thereby regulating the flow and direction of hydraulic oil to the steering cylinder to achieve the desired steering dynamics. This algorithm compensates for the system's inertia and damping characteristics through a feedforward loop, while simultaneously using PID feedback to correct tracking errors in real time. For trajectory planning, the system employs a ramp function algorithm for speed planning, ensuring a smooth, shock-free steering process while satisfying speed and acceleration constraints, achieving coordinated and synchronous control of multi-axis steering. Attached Figure Description

[0018] Figure 1 This is a system schematic diagram of the present invention.

[0019] Figure 2 This is a schematic diagram of the dual front axle steering mode of the present invention.

[0020] Figure 3 This is a schematic diagram of the dual rear axle steering mode of the present invention.

[0021] Figure 4 This is a schematic diagram of the four-bridge steering mode of the present invention.

[0022] Figure 5 This is a schematic diagram of the oblique mode of the present invention. Detailed Implementation

[0023] Specific embodiments of the present invention will now be described with reference to the accompanying drawings. Many details are described below to provide a comprehensive understanding of the invention; however, those skilled in the art will not need these details to implement the invention. Well-known components, methods, and processes will not be described in detail below.

[0024] This embodiment discloses a distributed cooperative control method for hydraulic braking and steering of a port heavy-duty IGV, based on a port heavy-duty IGV hydraulic braking and steering control system. The port heavy-duty IGV is a four-axle IGV. (Refer to...) Figure 1 The port heavy-duty IGV hydraulic braking and steering control system includes a hydraulic station 8, a steering system, a braking system, and a controller.

[0025] Reference Figure 1 The hydraulic station 8 consists of two relief valves 9, a solenoid ball valve 10, a hand-cranked pump 11, a high-pressure motor pump 12, and a low-pressure emergency motor pump 13. During normal operation, the high-pressure motor pump 12 operates. When the high-pressure circuit fails, the low-pressure emergency motor pump 13 can operate briefly to allow the vehicle to pull over. If both the high-pressure motor pump 12 and the low-pressure emergency motor pump 13 fail, the vehicle can be towed using the hand-cranked pump 11 in conjunction with the system's manual operating valve to release the vehicle's parking brake. The solenoid ball valve 10 of the hydraulic station is mainly used for unloading the hydraulic system. When the accumulator reaches the target pressure, the solenoid ball valve 10 is de-energized to unload, and the motor pump operates under low load. When the accumulator pressure falls below the set pressure, the solenoid ball valve 10 is energized to build up pressure and replenish the accumulator. The relief valve 9 of the hydraulic station is used to design the system's maximum pressure.

[0026] Reference Figure 1 The steering system includes steering cylinders, steering cylinder pressure sensors, steering accumulators, and steering valve assemblies. The steering cylinders are designated as follows: left-side steering cylinder 3-1 (first axle), right-side steering cylinder 3-2 (first axle), left-side steering cylinder 3-3 (second axle), right-side steering cylinder 3-4 (second axle), left-side steering cylinder 3-5 (third axle), right-side steering cylinder 3-6 (third axle), left-side steering cylinder 3-7 (fourth axle), and right-side steering cylinder 3-8 (fourth axle). The steering cylinder pressure sensors are designated as follows: pressure sensor 4-9 (first axle), pressure sensor 4-10 (second axle), pressure sensor 4-11 (third axle), and pressure sensor 4-12 (fourth axle). The steering accumulators are designated as follows: steering accumulator 5-3 (first axle), steering accumulator 5-4 (second axle), steering accumulator 5-5 (third axle), and steering accumulator 5-6 (fourth axle). The steering valve assemblies are designated as follows: steering valve assembly 6-3 (first axle), steering valve assembly 6-4 (second axle), steering valve assembly 6-5 (third axle), and steering valve assembly 6-6 (fourth axle). The steering valve group 6-3 (one axle), steering valve group 6-4 (two axles), steering valve group 6-5 (three axles), and steering valve group 6-6 (four axles) are each equipped with a steering proportion valve.

[0027] Reference Figure 1The braking system includes brake cylinders, brake accumulators, brake cylinder sensors, brake accumulator sensors, and brake valve assemblies. The brake cylinders include: left-side service brake cylinder 1-1 (one axle), right-side service brake cylinder 1-2 (one axle), left-side service brake cylinder 1-3 (two axles), right-side service brake cylinder 1-4 (two axles), left-side service brake cylinder 1-5 (three axles), right-side service brake cylinder 1-6 (three axles), left-side service brake cylinder 1-7 (four axles), right-side service brake cylinder 1-8 (four axles), left-side parking brake cylinder 2-1 (two axles), and right-side parking brake cylinder 2-2 (three axles). The brake accumulators include: brake accumulators 5-1 (one and two axles) and 5-2 (three and four axles). The brake valve assemblies include: brake valve assembly 6-1 (one and two axles) and 6-2 (three and four axles). The brake accumulator sensors include: pressure sensor 4-1 (one and two axles) and pressure sensor 4-5 (three and four axles). The brake cylinder sensors include pressure sensor 4-2 for the single-axle service brake cylinder, pressure sensor 4-3 for the double-axle service brake cylinder, pressure sensor 4-6 for the triple-axle service brake cylinder, pressure sensor 4-7 for the quadruple-axle service brake cylinder, pressure sensor 4-4 for the double-axle parking brake, and pressure sensor 4-8 for the triple-axle parking brake cylinder.

[0028] The controller of this invention adopts a three-controller parallel topology, all of which use EHPSBU controllers, and are respectively numbered as EHPSBU01 controller, EHPSBU02 controller and EHPSBU03 controller.

[0029] The EHPSBU01 controller, as the coordination and control unit for braking and hydraulic power, constitutes the core of the system's braking control. This controller implements a complete braking function system, including dynamic pressure regulation of the service brake and safety control of the parking brake. In the communication architecture, EHPSBU01 acts as a gateway between the internal and external CAN networks. On one hand, it forwards hydraulic power control commands to EHPSBU02 via the internal CAN bus; on the other hand, it converts the hydraulic power system status information fed back by EHPSBU02 and forwards it to the external CAN bus, achieving vehicle-level information exchange and status monitoring.

[0030] The EHPSBU02 controller serves as both a steering actuation and hydraulic power monitoring unit, possessing dual functional positioning. In terms of steering control, this controller is responsible for the steering actuation control of the first and second axles of the port heavy-duty IGV, achieving precise control of the steering angle and real-time feedback of the steering system status. Regarding hydraulic power management, the EHPSBU02 executes the start-stop control logic of the hydraulic power system, while simultaneously monitoring key parameters such as system temperature and pressure in real time, implementing a multi-layered system protection mechanism to ensure the safe and reliable operation of the hydraulic system.

[0031] The EHPSBU03 controller, a dedicated steering control unit, focuses on steering execution control for both three and four axes. This controller achieves high-precision steering angle tracking control and, through a status monitoring mechanism, collects and reports the operating parameters of the steering system in real time, providing accurate feedback information to the upper-level control system to ensure the coordination and precision of multi-axis steering.

[0032] During steering control, the controller, as the core processing unit of the system, calculates the target steering angle and target angular velocity in real time based on the steering commands issued by the vehicle controller. Based on this, the controller uses a composite control algorithm combining feedforward and PID feedback to generate corresponding control signals, precisely driving the opening of the steering proportional valve, thereby regulating the flow and direction of hydraulic oil to the steering cylinder to achieve the desired steering dynamics characteristics. This algorithm compensates for the system's inertia and damping characteristics through the feedforward stage, while simultaneously using the PID feedback stage to correct tracking errors in real time.

[0033] To construct a high-precision closed-loop feedback control system, angle sensors mounted on the kingpin of the steering knuckle of the axle detect and provide feedback on the actual steering angle of each axle in real time. The controller compares the feedback steering angle with the target steering angle in real time, calculates the deviation using a PID control algorithm, and dynamically adjusts the control signal of the proportional valve until the actual steering angle precisely converges to the target steering angle, thereby achieving precise servo control of the steering process. For trajectory planning, the system employs a ramp function algorithm for speed planning, ensuring a smooth and shock-free steering process while satisfying speed and acceleration constraints, achieving coordinated and synchronous control of multi-axle steering.

[0034] The distributed cooperative control method for hydraulic braking and steering of heavy-duty IGVs in ports according to the present invention specifically includes:

[0035] 1. Define the following objective function:

[0036] (1)

[0037] in, i For each axis, Maximum speed for each axis, t For the motion time of all axes, yes i Axis control input, T This is the time required for all axes to reach the target steering angle;

[0038] 2. The angle of each axis is The initial angle is The acceleration of each axis is The motion of each axis is divided into three stages: acceleration, uniform motion, and deceleration. Assuming that the magnitudes of acceleration and deceleration are the same, the equations of motion are as follows:

[0039] (1) Acceleration phase : (2); where, (3);

[0040] (2) Uniform speed stage : (4); among which, (5);

[0041] (3) Deceleration phase : (6);

[0042] Total motion time for each axis It can be represented as: (7).

[0043] To synchronize the motion of all axes, the maximum motion time is taken as the synchronization time: (8).

[0044] To ensure that all axes complete their motion within time T, the maximum speed of each axis is adjusted as follows: (9).

[0045] The dynamic equations for each axis are: (10); among which, yes i Moment of inertia of the shaft yes i The damping coefficient of the shaft, yes i Axis control input, yes i The actual angular acceleration of the axis, yes i The actual angular velocity of the shaft;

[0046] 3. Design the control input using a PID controller or feedforward control strategy. :

[0047] (11)

[0048] in, For the desired angle position, For the desired angular velocity, For the desired angular acceleration, For proportional gain, This is the differential gain.

[0049] Figures 2 to 5The multi-mode steering capability of the port heavy-duty IGV is revealed, stemming from the independent, distributed control and coordination of the steering angles of each axle. In terms of control strategy, the system is based on the Ackermann steering principle, precisely calculating the target steering angle of each wheel to ensure they move around a common instantaneous center of rotation during steering, thereby optimizing tire dynamics. Based on this architecture, the system can achieve multiple steering modes, including: a dual-front axle steering mode suitable for high-speed cruising, a dual-rear axle steering mode suitable for operation in confined spaces, an all-wheel steering mode with a very small turning radius, and a diagonal mode for lateral translation.

[0050] Figures 2 to 5 In this diagram, I, II, III, and IV refer to the first, second, third, and fourth axles of a port heavy-duty IGV, respectively. Using the first axle as the reference point for the vehicle's front end, and observing along the vehicle's forward direction, the steering angles are defined as positive to the left and negative to the right. The system only needs to provide the interior angle of the first axle. This will automatically determine the interior angles of axes two through four. The interior angles of axes one through four are set as follows: , , , The formulas for calculating the interior angles of each mode are as follows:

[0051] Dual front axle steering mode: (12);

[0052] Dual rear axle steering mode: (13);

[0053] Four-axle steering mode: (14);

[0054] Diagonal mode: (15);

[0055] In the formulas for calculating the interior angle of each steering mode above, L1 refers to the wheelbase between the first and second axles; L2 refers to the wheelbase between the second and third axles; and L3 refers to the wheelbase between the third and fourth axles.

[0056] In terms of safety, the relief valve, acting as a pressure safety valve for the system, has a set maximum pressure threshold. When the circuit pressure suddenly rises due to abnormal operating conditions, the relief valve immediately opens to relieve pressure, limiting it within a safe range and protecting precision components such as the steering proportional valve from overpressure damage. In the circuit design, a check valve is installed between the steering accumulator and the steering proportional valve to ensure unidirectional hydraulic energy supply. This effectively locks the high pressure from the steering accumulator to the steering proportional valve, maintaining stable inlet pressure even when the main pump is interrupted or pressure fluctuates, ensuring rapid steering response. The steering accumulator, as an auxiliary power source, stores and releases hydraulic potential energy. When the system's required flow exceeds the main pump's supply capacity or the main pump experiences a temporary failure, the steering accumulator quickly replenishes energy, thereby improving the system's dynamic response, absorbing pressure pulsations, and serving as an emergency power source to enhance system robustness. The steering cylinder pressure sensor is responsible for real-time monitoring of the circuit pressure.

[0057] Based on the theory of optimal braking force distribution, the braking system adopts an integrated hydraulic valve group architecture. Both the primary and secondary brake valve groups integrate two pressure-reducing proportional valves, two solenoid valves, one solenoid valve with a manual operation knob, and a check valve. The braking system achieves optimal braking force distribution by solving a multi-objective optimization function centered on the sum of squares of braking force, while satisfying the total braking force requirements, actuator physical limitations, and dynamic response constraints.

[0058] In parking brake control mode, the system precisely opens the parking brake valve by energizing the solenoid valve, thereby releasing the mechanical constraint of the parking brake caliper. In service brake control mode, the system achieves precise pressure control based on the braking percentage command by adjusting the opening of the pressure reducing proportional valve, and constructs a closed-loop control circuit with pressure sensor feedback at its core. Simultaneously, it coordinates the opening and closing states of the solenoid valve to execute the service brake function. This control strategy, through the coordinated operation of the two braking modes, ensures the reliability and accuracy of the system under different operating conditions.

[0059] The one-way valve in the braking system ensures unidirectional flow of hydraulic energy, maintaining a high-pressure state from the steering accumulator to the steering proportioning valve, thus providing a stable hydraulic power source for the steering system. The steering accumulator, as an auxiliary hydraulic energy source, compensates for flow rate under peak system demand conditions, significantly improving the system's dynamic response performance. Pressure sensors monitor the pressure parameters of each circuit in real time, providing crucial status feedback information for optimized braking force distribution, ensuring the system achieves optimal control performance under multiple constraints.

[0060] To ensure the most effective control of braking force, the objective function is designed as follows:

[0061] (16);

[0062] in, For the first Axis in t Braking force at all times for t Total braking force requirement at any moment For the first i Maximum braking force limit of the shaft, For the first i Limitation on the rate of change of braking force on the shaft. , These are the weighting coefficients. t f This represents the total braking time.

[0063] The above are merely specific embodiments of the present invention, but the design concept of the present invention is not limited thereto. Any non-substantial modifications made to the present invention using this concept shall be considered as infringing upon the protection scope of the present invention.

Claims

1. A distributed cooperative control method for hydraulic braking and steering of a port heavy-duty IGV, based on a port heavy-duty IGV hydraulic braking and steering control system, including a hydraulic station, a steering system, a braking system, and a controller. The port heavy-duty IGV is a four-axle IGV. An angle sensor on the kingpin of the steering knuckle of the axle detects and feeds back the steering angle of each axle in real time. The controller compares the feedback steering angle with the target steering angle in real time, calculates the deviation through a PID control algorithm, and dynamically adjusts the control signal to the steering proportional valve until the actual angle precisely converges to the target steering angle, thereby achieving precise servo control of the steering process; characterized in that... Specifically, the process includes the following: I. Define the following objective function: ,in, For each axis, Maximum speed for each axis, t For the motion time of all axes, yes i Axis control input, T This is the time required for all axes to reach the target steering angle; II. Velocity planning using the ramp function algorithm: (1) Divide the motion of each axis into three stages: acceleration, constant speed, and deceleration. Assume that the magnitudes of acceleration and deceleration are the same, and the angle of each axis is . The initial angle is The acceleration of each axis is (2) Calculate the total motion time for each axis. (3) Take the maximum motion time as the synchronization time: (4) Adjust the maximum speed of each axis to (5) The dynamic equations for each axis are: ;in, yes i Moment of inertia of the shaft yes i The damping coefficient of the shaft, yes i The actual angular acceleration of the axis, yes i The actual angular velocity of the shaft; III. Design using a PID controller and feedforward control strategy i Axis control input : ,in, For the desired angle position, For the desired angular velocity, For the desired angular acceleration, For proportional gain, This is the differential gain.

2. The distributed cooperative control method for hydraulic braking and steering of a port heavy-duty IGV as described in claim 1, characterized in that: The motion equations in step (2) of the speed planning in step two are as follows: ① Acceleration phase : ,in, ② Uniform speed stage : ,in, ③ Deceleration phase : .

3. The distributed cooperative control method for hydraulic braking and steering of a port heavy-duty IGV as described in claim 2, characterized in that: The total motion time of each axis in step two Represented as: Maximum speed for: , t i1 Let be the acceleration time along the i-axis.

4. The distributed cooperative control method for hydraulic braking and steering of a port heavy-duty IGV as described in claim 1, characterized in that: The port heavy-duty IGV hydraulic brake-steering control system implements the following steering modes, including: dual front axle steering mode suitable for high-speed cruising, dual rear axle steering mode suitable for operation in narrow spaces, and all-wheel steering mode with a very small turning radius.

5. The distributed cooperative control method for hydraulic braking and steering of a port heavy-duty IGV as described in claim 4, characterized in that: The interior angles of the first to fourth axles of the heavy-duty IGV in the port are set as follows: , , , The formulas for calculating the interior angles of each mode are as follows: Dual front axle steering mode: Dual rear axle steering mode: Four-axle steering mode: Diagonal mode: ; In the formulas for calculating the interior angle of each steering mode above, L1 is the interior angle of the first axis given by the system; L2 is the wheelbase between the first and second axes; L3 is the wheelbase between the second and third axes; L4 is the wheelbase between the third and fourth axes.

6. The distributed cooperative control method for hydraulic braking and steering of a port heavy-duty IGV as described in claim 3, characterized in that, To ensure the most effective control of braking force, the objective function is designed as follows: ;in, For the first Axis in t Braking force at all times for t Total braking force requirement at any moment For the first i Maximum braking force limit of the shaft, For the first i Limitation on the rate of change of braking force on the shaft. , These are the weighting coefficients. t f This represents the total braking time.

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