Active-passive wave compensation hydraulic control system applied to deep sea hoisting

Abstract

The invention discloses an active-passive wave compensation hydraulic control system applied to deep sea hoisting, which comprises a speed and active compensation control module, a tension and passive compensation control module, an energy accumulator, a power station and a brake control module, the tension and passive compensation control module adopts a variable and constant displacement motor series connection structure, drives the winding drum to perform tension control in the tension mode through a gear box according to the two-stage control principle, and provides balance torque in the speed mode; according to the power station module, a servo motor drives a metering pump to supply oil, braking is controlled by a reversing valve to start and stop, and the system coordinates static load and dynamic compensation based on speed and force load sensor feedback and adapts to complex sea conditions. The device has the characteristics of fast response, high tension control precision and low energy consumption.

Description

Technical Field

[0001] This invention relates to the field of hydraulic control technology, and more specifically, to the field of hydraulic control technology for deep-sea engineering equipment. Background Technology

[0002] During deep-sea lifting operations, the complex and ever-changing sea conditions and the relative displacement of the hull caused by waves can lead to periodic up-and-down oscillations of the lifting load, seriously affecting the safety and accuracy of the operation. To reduce the impact of wave disturbance on lifting operations, the industry generally adopts wave compensation technology, which uses hydraulic drive and control systems to adjust the output speed or tension of the drum to maintain the relative stability of the lifted position.

[0003] In existing technologies, active compensation systems offer fast response speeds and high control precision, but they require high power system capacity and servo control performance, resulting in large overall installed power and high energy consumption. Passive compensation systems rely on accumulators for energy absorption and release, offering advantages such as simple structure and low energy consumption, but their compensation performance is easily affected by accumulator characteristics, making them difficult to adapt to complex sea conditions. To balance high response and energy efficiency, a hybrid compensation mode combining active and passive methods is gradually becoming the development trend for deep-sea lifting equipment.

[0004] However, the existing hybrid wave compensation system still has the following problems: (1) The control logic of the active compensation link is complex, the speed control is difficult, and the system stability and adaptability are insufficient; (2) The energy absorption and release efficiency of the passive compensation system is not high, and there is a large pressure drop during the connection process between the accumulator and the system, which affects the compensation performance; (3) The piston seal of the accumulator is prone to wear or failure, which reduces the reliability of the system; (4) The overall system power demand is too high, and it is difficult to balance rapid response and energy consumption optimization. Summary of the Invention

[0005] To address the aforementioned technical problems, this invention proposes a hydraulic control system combining active and passive wave compensation. By setting the secondary control loop as passive compensation control and employing a piston accumulator to optimize energy management, while introducing a proportional servo valve pressure compensation control motor module, collaborative control of speed closed loop, constant tension open loop, and constant tension closed loop is achieved. This improves compensation performance, reduces control complexity, and enhances system energy efficiency and reliability.

[0006] To achieve the above objectives, the present invention provides the following technical solution: An active-passive wave-compensated hydraulic control system for deep-sea lifting, characterized in that the system comprises:

[0007] The speed and active compensation control module includes an active compensation motor, whose inlet is connected to the system's high-pressure oil circuit via a proportional servo valve, and whose inlet pressure is stabilized via an oil source selection passage and a pressure compensation passage. The output of the active compensation motor drives a drum via gearbox III to provide dynamic compensation torque.

[0008] The tension and passive compensation control module includes a variable displacement motor and a fixed displacement motor connected in series. Its inlet is connected to the high-pressure oil circuit of the system and the pilot pressure is kept stable through the pilot stabilization passage. The output of the tension and passive compensation control module drives the drum through the gearbox to output static load compensation torque.

[0009] The accumulator module includes an accumulator that is selectively connected to the high-pressure oil circuit of the system via an isolation valve, and in the event of a power station failure, the piston accumulator switches via a reversing valve IV to provide an emergency oil source to the system pilot control circuit.

[0010] The power station module includes a fixed displacement pump driven by a servo motor that supplies oil to the system's high-pressure oil circuit.

[0011] The tension and passive compensation control module and the speed and active compensation control module are both mechanically connected to the drum. The encoder is set at the shaft end, which belongs to a fixed-displacement motor or variable-displacement motor connected to the drum gear transmission. It is used to characterize the drum rotation speed and the detection signal of the force load sensor of the wire rope set on the drum, and to coordinate and control the static load compensation torque and dynamic compensation torque output by the two modules.

[0012] Furthermore, the variable motor swashplate angle is controlled by a closed-loop position control system consisting of a servo valve and a swashplate angle sensor to achieve precise adjustment of the static load compensation torque.

[0013] Furthermore, the variable displacement motor and the fixed displacement motor are connected in series with a displacement ratio of 1:1, thereby forming a static load compensation unit capable of outputting torque from zero to maximum.

[0014] Furthermore, the high-pressure inlet of the tension and passive compensation control module is characterized in that the pressure circuit is controlled to open / close via a logic valve, and the opening or closing action of the logic valve controls whether the high-pressure inlet is connected to the high-pressure oil circuit of the system.

[0015] Furthermore, the electromagnetic reversing valve V of the speed and active compensation control module can switch the active compensation motor to active control mode when energized, and switch it to low differential pressure floating mode when de-energized, in order to adapt to the operation requirements of constant tension control or active speed control, and the pilot circuit flushing and oil replenishment is provided by the one-way valve XXIX.

[0016] Furthermore, the speed and active compensation control module enables the active compensation motor to automatically select the higher pressure oil source from the two parallel one-way valves via a parallel one-way oil source selection path.

[0017] Furthermore, the speed and active compensation control module includes a pressure compensator composed of shuttle valve I and pressure reducing valve II, which is used to compensate the oil supply pressure of the active compensation motor. A bidirectional balance valve is provided on the oil return side of the active compensation motor to adjust its motion state during operation, thereby ensuring the controllability and stability of the active compensation motor operation.

[0018] Furthermore, it also includes a brake control module, which includes a brake disposed in the drum drive link, used to release or brake by supplying or releasing control oil pressure to the brake actuator when the torque balance condition is met.

[0019] Furthermore, before the brake is released, the brake control module outputs a pre-balancing torque through the variable motor and the fixed motor. After triggering the torque balancing proximity switch, the reversing valve XII realizes the release or braking of the brake.

[0020] Furthermore, when the piston accumulator outputs control oil as an emergency oil source, it limits the flow rate through a one-way speed regulating valve. When the power station module outputs oil, it works in conjunction with the pulsation absorption accumulator to provide or absorb instantaneous large flow oil sources and reduce pressure pulsations.

[0021] The technical solution proposed in this invention has the following effects:

[0022] 1. High efficiency and energy saving, significantly reducing system installed power and operating costs.

[0023] The system utilizes a piston-type accumulator to absorb and store energy during the wave's ascent phase and release it back to the system during the wave's descent phase, through the coordinated operation of the tension and passive compensation control module and the accumulator module. This "peak shaving and valley filling" effect means that the system's power station does not need to be designed according to the peak power during operation, but only needs to meet the average power requirement. This significantly reduces the installed power of core power components such as servo motors and hydraulic pumps, thereby reducing initial investment and long-term operating energy consumption.

[0024] 2. High control precision, fast response speed, and superior dynamic performance.

[0025] The system achieves an organic combination of active and passive compensation. Passive compensation bears most of the steady-state load, while active compensation provides dynamic torque quickly and accurately through a high-response module composed of a proportional servo valve, a pressure compensator, and a bidirectional balancing valve assembly. This design ensures the system's rapid tracking of wave motion while effectively overcoming the control lag problem caused by the large inertia of purely electrical wave compensation. Simultaneously, the introduction of real-time correction of the wire rope winding radius and friction torque compensation based on the measured intermediate hysteresis curve further improves the control accuracy of tension and speed.

[0026] 3. Safe and reliable, with multi-layered redundancy and emergency response mechanisms.

[0027] The system is designed with the high risks of deep-sea operations in mind, incorporating multiple safety measures. Key components such as the swashplate control servo valve and encoder of the variable displacement motor are redundantly designed to ensure system functionality is maintained in the event of a single component failure. The main pump unit of the power station is also redundantly configured and can quickly isolate faulty units through logic valves and directional valves. The unique brake release logic (requiring confirmation from the torque balance proximity switch) and the emergency fuel supply provided by the accumulator module in the event of power failure together constitute a comprehensive safety system from fault prevention and fault-tolerant operation to emergency support.

[0028] 4. System optimization to effectively balance complexity and cost.

[0029] Through ingenious configuration design, the system optimizes cost and complexity while ensuring performance. It employs a parallel configuration of variable displacement and fixed displacement motors in a 1:1 displacement ratio, achieving a wide range of stepless displacement adjustment in a relatively economical way. Simultaneously, the speed control function is separated from the traditional complex two-stage control system and handled by an independent speed and active compensation control module. This significantly reduces the difficulty and complexity of the system's core control algorithm, improving the system's maintainability and overall reliability. Attached Figure Description

[0030] Figure 1 System Schematic Diagram

[0031] Figure 2 Power station module schematic diagram

[0032] Figure 3 Schematic diagram of energy storage module

[0033] Figure 4 Schematic diagram of tension and passive compensation control module

[0034] Figure 5 Schematic diagram of speed and active compensation control module

[0035] Figure 6 Constant tension open-loop control logic diagram

[0036] Figure 7 Constant tension closed-loop control logic diagram

[0037] Figure 8 Variable motor swashplate angle-load relationship curve

[0038] Figure 9 Drum speed closed-loop control logic diagram

[0039] Figure 10 System overall structure diagram

[0040] Label Explanation 1: Metering pump I, 1': Metering pump II, 2: Servo motor I, 2': Servo motor II, 3: Relief valve I, 3': Relief valve II, 4: Check valve I, 4': Check valve II, 5: Logic valve I, 5': Logic valve II, 6: Directional control valve I, 6': Directional control valve II, 7: Motor, 8: Low-pressure pump, 9: Check valve III, 10: Cooler, 11: Check valve IV, 12: Check valve V, 13: Oil tank, 21: Piston accumulator, 22: Magnetostrictive sensor, 23: Safety valve, 24: Pressure sensor I, 25: Gas cylinder, 31: Directional control valve core I, 32: Speed ​​control valve, 33: Directional control valve III, 34: Pressure sensor II, 35: Isolation valve, 36: Check valve VI, 37: Check valve VII, 38: Reversing seat valve; 39: Relief valve III; 40: Reversing valve IV; 41: Relief valve IV; 42: Reversing valve VI; 43: Pressure monitoring sensor; 44: Check valve VIII; 45: One-way speed control valve; 46: Pressure reducing valve I; 47: Check valve IX; 50: Pressure sensor III; 51: Pressure sensor IV; 52: Parallel accumulator I; 52': Parallel accumulator II; 52'': Parallel accumulator III; 53: Pulse absorption accumulator I; 53': Pulse absorption accumulator II; 54: Relief valve V; 55: Relief valve VI; 80: Gearbox I; 80': Gearbox II; 81: Slewing bearing; 82: Drum; 83: Gearbox III; 100: Reversing valve VII; 100': Reversing valve VIII. 101: Logic Valve III, 101': Logic Valve IV, 102: Pressure Sensor V, 102': Pressure Sensor VI, 103: Check Valve X, 103': Check Valve XI, 104: Check Valve XII, 104': Check Valve XIII, 105: Check Valve XIV, 105': Check Valve XV, 106: Relief Valve VII, 106': Relief Valve VIII, 107: Torque Balance Proximity Switch I, 107': Torque Balance Proximity Switch II, 108: Brake I, 108': Brake II, 109: Check Valve XVI, 109': Check Valve XVII, 110: Variable Motor I, 110': Variable Motor II, 111: Swashplate Angle Sensor I, 111': Swashplate Angle Sensor II, 120: Solenoid Directional Valve I, 120': Solenoid directional valve II, 121: Directional valve spool II, 121': Directional valve spool III, 122: Directional valve spool IV, 122': Directional valve spool V, 123: Main servo valve I, 123': Main servo valve II, 124: Standby servo valve I, 124': Standby servo valve II, 125: Directional valve spool VI, 125': Directional valve spool VII,126: Directional control valve spool VIII, 126': Directional control valve spool IX, 127: Solenoid directional control valve III, 127': Solenoid directional control valve IV, 130: Directional control valve IX, 130': Directional control valve X, 131: Logic valve V, 131': Logic valve VI, 132: Check valve XVIII, 132': Check valve XIX, 133: Check valve XX, 133': Check valve XXI, 134: Check valve XXII, 134': Check valve XXIII, 135: Relief valve IX, 135': Relief valve X, 136: Fixed displacement motor I, 136': Fixed displacement motor II, 137: Encoder I, 137': Encoder II, 138: Check valve XXV, 138': One-way valve XXVI, 140: Oil accumulator, 200: Proportional servo valve, 201: Shuttle valve I, 202: Pressure reducing valve II, 203: Two-way balance valve I, 204: Pressure sensor VIII, 205: Pilot-operated relief valve I, 206: Shuttle valve II, 210: Two-way balance valve II, 211: Pilot-operated relief valve II, 212: Pressure sensor IX, 213: One-way valve XXIV, 214: Solenoid directional valve V, 215: Relief valve XI, 220: One-way valve XXVII, 221: One-way valve XXVIII, 222: One-way valve XXIX, 223: Active compensation motor, 300: Pressure reducing valve III, 301: Directional valve XII, 302: Pressure sensor VII, 400: Force load sensor. Detailed Implementation

[0041] The technical solutions of the embodiments of the present invention will be clearly and completely described below with reference to the accompanying drawings. Obviously, the described embodiments are only some embodiments of the present invention, and not all embodiments. Based on the embodiments of the present invention, all other embodiments obtained by those skilled in the art without creative effort are within the scope of protection of the present invention.

[0042] This invention provides an active-passive wave-compensated hydraulic control system for deep-sea lifting operations. For example... Figure 1 As shown in the system schematic diagram, the system generally includes a tension and passive compensation control module, a speed and active compensation control module, a brake control module, an accumulator module, and a power station module. These modules work together to drive and control the transmission mechanism connected to the drum 82 (such as...). Figure 10(As shown in the overall system structure diagram), including gearboxes 80, 80', 83 and slewing bearing 81, to achieve closed-loop speed control, open-loop constant tension control, and closed-loop constant tension control of the drum 82. The tension and passive compensation control module and the speed and active compensation control module are both mechanically connected to the drum 82. Encoders 137 and 137' are located at the shaft ends of the quantitative motors 136 and 136', which are connected to the gear transmission of the drum 82. They are used to characterize the rotational speed of the drum 82, and the detection signal from the force load sensor 400 on the wire rope of the drum 82 is used to coordinate the static load compensation torque and dynamic compensation torque output by the two modules, thereby achieving passive-active hybrid wave compensation of the drum 82 during deep-sea hoisting operations.

[0043] The tension and passive compensation control module is the core load-bearing unit of the system, and its principle is detailed in [link to relevant documentation]. Figure 4 (Schematic diagram of tension and passive compensation control module). This module uses variable displacement motors 110 and 110' connected in series with fixed displacement motors 136 and 136', whose inlets are connected to the high-pressure oil circuit of the system ( Figure 4The solid red line in the middle represents the high-pressure oil circuit of the system, which is connected and maintains pilot pressure stability through a pilot stabilization path. The output of the tension and passive compensation control module drives the drum 82 via gearboxes 80 and 80' to output static load compensation torque. In this module, the displacement ratio of variable motors 110 and 110' to fixed displacement motors 136 and 136' is 1:1, thus forming a static load compensation unit from zero to maximum torque. To ensure uniform force distribution, variable motors 110 and 110' are of the same model and control swashplate angles to be consistent, as are fixed displacement motors 136 and 136'. The swashplate angle of variable motors 110 and 110' is controlled by a closed-loop position control system consisting of servo valves 123 and 123' and swashplate angle sensors 111 and 111' to achieve precise adjustment of static load compensation torque. Variable motors 110 and 110' are equipped with swashplate angle sensors 111 and 111' for real-time feedback of swashplate position. This module uses main servo valves 123 and 123' to perform precise closed-loop position control of the swashplate angle of variable motors 110 and 110'. The main servo valves 123 and 123' achieve closed-loop position control by energizing solenoid directional valves 120 and 120' to control the reversing of directional valve cores 121, 121', 122, and 122'. To improve system reliability, backup servo valves 124 and 124' are provided. When the main servo valves 123 and 123' fail, the solenoid directional valves 120 and 120' are de-energized, causing the directional valve cores 121, 121', 122, and 122' to reset. At the same time, the solenoid directional valves 127 and 127' are energized, thereby driving the directional valve cores 125, 125', 126, and 126' to operate, switching the backup servo valves 124 and 124' into the control loop. The backup servo valves 124 and 124' receive the same control command signals as the main servo valves 123 and 123', ensuring that the control mode of the variable motors 110 and 110' remains consistent before and after the switch, thus maintaining continuous and stable system operation. Meanwhile, considering the large flow rate changes required for the control process of the main servo valves 123 and 123', and the potential for pressure drop due to untimely oil replenishment when simply drawing oil from the system's piston accumulator 21 via pipelines, an oil replenishment accumulator 140 was added near the main servo valves 123 and 123'. After all valve groups fail, the variable motors 110 and 110' automatically return to their maximum displacement position under the action of the circuit pressure difference and their own reset structure, thus entering a passive load state. At this time, the basic pilot pressure required for the pilot control circuit of the variable motors 110 and 110' is provided by the accumulator module, and the pressure of its emergency pilot oil source is monitored in real time by the pressure monitoring sensor 43 installed on the pilot oil line to determine whether the emergency oil supply status meets the system's safe operation requirements.

[0044] like Figure 1As shown, the red line area represents the system's high-pressure oil circuit, and the blue line area represents the system's low-pressure oil circuit. The high-pressure inlet of the tension and passive compensation control module is controlled to connect or disconnect the pressure circuit through logic valves 101, 101' and 131, 131', respectively. Logic valves 101, 101' are used to control whether the high-pressure inlet corresponding to the variable displacement motor is connected to the system's high-pressure oil circuit, and logic valves 131, 131' are used to control whether the high-pressure inlet corresponding to the fixed displacement motor is connected to the system's high-pressure oil circuit. The logic valves 101, 101' are enabled by pilot pressure provided by directional valves 100, 100', and the logic valves 131, 131' are enabled by pilot pressure provided by directional valves 130, 130'. To ensure stable pilot control pressure, multiple sets of check valves are installed in the oil circuit, including check valves 103, 103', 104, 104', 105, and 105' (serving the variable displacement motor circuit) and check valves 132, 132', 133, 133', 134, and 134' (serving the fixed displacement motor circuit). Relief valves 106, 106' and 135, 135' are used to protect the corresponding motors from exceeding their design pressure. Furthermore, the high-pressure sides of the variable displacement motors 110, 110' and the fixed displacement motors 136, 136' are replenished with oil via check valves 109 and 109'. Gearboxes 80 and 80' are equipped with brakes 108 and 108', as well as torque balance proximity switches 107 and 107'. The static load torque output by this module is provided by the motor drive torque Tm1, the value of which is determined according to the load balance equation Tm1 + Tm2 = Tls + J × α + Tf, where Tls is the equivalent torque of the static load, J is the current moment of inertia of the system (by adding an extra moment of inertia to the gearbox, ensuring that Jmax / Jmin ≤ 1.5), α is the acceleration required by the system (limiting the acceleration time to no more than 0.2 seconds), and Tf is the changing friction torque (the magnitude and direction of Tf will change when the motion accelerates and decelerates). The motor drive torque Tn is calculated using the formula Tn = Vn × (p1 - p2) × ηhm / (20 × π) = βn × kn × (p1 - p2) × ηhm / (20 × π), where Vn is the motor displacement, βn is the motor swashplate angle, kn is the amplification factor related to the motor model, p1 is the high-pressure port pressure of the motor (detected by pressure sensors 102 and 102'), p2 is the low-pressure port pressure of the motor (detected by pressure sensor IV 51), and ηhm is the overall mechanical-hydraulic efficiency. To improve control accuracy and simplicity, the system utilizes the measured values... Figure 8The intermediate hysteresis curve (code 450) in the (variable motor swashplate angle-load relationship curve) is used for compensation, ensuring that Tm1 ≈ Tls + 0.5 × Tf. This guarantees that the drum remains in a safe state of floating but not rotating under passive compensation. To further improve the tension control accuracy on the wire rope, the system also incorporates wire rope winding radius data (via... Figure 10 The speed signals of encoders 137 and 137' are integrated to obtain the current number of wire rope layers, and then the radius is determined to correct the motor torque. Figure 10 The force load sensor 400 shown is used to detect the real-time tension of the wire rope (note the geometric relationship between the force applied by the wire rope and the force load sensor 400; if correction is needed, a correction formula must be introduced to ensure accurate measurement). Its signal is compared with the set value to adjust the control output. Figure 6 (Constant tension open-loop control logic diagram) and Figure 7 (Constant Tension Closed-Loop Control Logic Diagram) shows the control mode where, in constant tension control (including open-loop and closed-loop) mode, the solenoid directional valve V 214 is de-energized, causing the pilot-operated relief valves 205 and 211 to be in the minimum pressure setting state. This allows the active compensation motor 223 to achieve low differential pressure operating conditions and enter a floating state. In this state, the low-pressure side oil forms a feedback loop through the Y-type function of the proportional servo valve 200, and replenishes oil to the two chambers of the active compensation motor 223 via the unidirectional oil replenishment channels of the bidirectional balance valves 203 and 210. This ensures that the two chambers of the active compensation motor 223 always have the necessary oil supply conditions under floating conditions, preventing loop venting and improving system response stability. In constant tension control mode, the system provides both open-loop and closed-loop control strategies, such as... Figure 6 (Constant tension open-loop control logic diagram) and Figure 7(Constant tension closed-loop control logic diagram) is shown. In this mode, the electromagnetic directional valve V214 is de-energized, causing the pilot-operated relief valves 205 and 211 to be in the lowest pressure setting state, thereby enabling the active compensation motor 223 to form a low pressure differential working condition and enter a floating state. The tension control task is entirely undertaken by the tension and passive compensation modules. At this time, the low-pressure side oil forms a feedback loop through the Y-type function of the proportional servo valve 200, and replenishes oil to the two chambers of the active compensation motor through the one-way oil replenishment channel of the bidirectional balance valves 203 and 210, to ensure that it always has the necessary oil supply conditions in the floating condition, prevents the loop from running dry, and improves the system response stability. In open-loop control, the system directly calculates the required swashplate angle control value of the variable motors 110 and 110' based on the preset test experience value (reference curve). At this time, the brakes 108 and 108' should be in the raised state. In closed-loop control, the real-time tension of the wire rope is detected by the force load sensor 400. The main servo valves 123 and 123' adjust the swashplate angle of the variable motor in a closed loop based on the error signal between the detected value and the system's target value, until the actual tension stabilizes near the target value. The signal from the force load sensor is compared with the set value to adjust the control output. If the installation geometry needs correction, appropriate formulas should be introduced to ensure measurement accuracy. Furthermore, in both control modes, the system calculates the effective winding radius of the wire rope by integrating the drum speed signal fed back by encoders 137 and 137', and accordingly corrects the torque output of the variable motor in real time. This compensates for tension fluctuations caused by changes in the drum radius, ultimately ensuring the overall accuracy of tension control.

[0045] The speed and active compensation control module is responsible for providing the dynamic torque Tm2, ​​the principle of which is detailed in [link to relevant documentation]. Figure 5(Schematic diagram of the speed and active compensation control module). The core actuator of this module is the active compensation motor 223. The inlet of the active compensation motor 223 is connected to the high-pressure oil circuit of the system via the proportional servo valve 200, and its inlet pressure is stabilized through the oil source selection passage and the pressure compensation passage. The output of the active compensation motor 223 drives the drum 82 via the gearbox III 83 to provide dynamic compensation torque. The equivalent driving torque of this module is Tm2 = (J × α + Tf) / 0.7. Its core actuator is the active compensation motor 223. Its inlet flow is controlled by the proportional servo valve 200. The shuttle valve I 201 and the pressure reducing valve II 202 together constitute a pressure compensator to ensure the stability of flow control. The bidirectional balance valves 203 and 210 ensure the controllability of the movement of the active compensation motor 223. The pilot-operated relief valves 205 and 211 are used to protect the active compensation motor 223 from exceeding the design pressure. The electromagnetic directional valve V 214 of the speed and active compensation control module can switch the active compensation motor 223 to active control mode when energized and switch it to low differential pressure floating mode when de-energized to adapt to the operating requirements of constant tension control or active speed control. One-way valve XXIX 222 provides pilot circuit flushing and oil replenishment. Specifically, when the electromagnetic directional valve V 214 is de-energized, pilot oil is led to the relief valve XI 215 through shuttle valve II 206 and one-way valve XXIV 213, setting the pressure of pilot-operated relief valves 205 and 211 to the low set value of relief valve XI 215, thus placing the active compensation motor 223 in a low-pressure floating state. When the electromagnetic directional valve V 214 is energized, the pilot-operated relief valves 205 and 211 return to their own spring's higher set value, and the active compensation motor 223 enters active control mode. The one-way valve XXIV 213 can prevent oil from flowing back into the shuttle valve II 206 under this condition.

[0046] The speed and active compensation control module, through a parallel unidirectional oil source selection path, enables the active compensation motor 223 to automatically select the higher pressure oil source from the two parallel unidirectional valves 220 and 221 connected to it, thus ensuring a stable oil supply pressure for the active compensation motor 223 and the proportional servo valve 200 under different operating conditions. Specifically, unidirectional valves 220 and 221 select the high-pressure oil source for the proportional servo valve 200 to avoid insufficient or fluctuating oil supply pressure affecting control accuracy; unidirectional valve XXIX 222 provides flushing flow to this part of the circuit to prevent oil circuit venting from causing slow system response. Figure 9As shown, in the speed and active compensation control module, pressure sensors 212 and 204 are respectively installed in the two chambers of the active compensation motor 223 to monitor the pressure in the two chambers of the active compensation motor 223 in real time. The pressure difference is used to characterize the magnitude and direction of the driving torque currently output by the active compensation motor 223. The control system calculates the pressure difference between the two chambers of the active compensation motor 223 based on the monitoring values ​​of pressure sensors 212 and 204, and uses this pressure difference as a criterion for the active compensation torque in the control logic judgment. The speed closed-loop control logic of this module is as follows: Figure 9 (The closed-loop control logic diagram for drum speed is shown). Under this control logic, the system first determines the state of brakes 108 and 108': when the brake is not released, the brake release process is executed first, i.e., according to... Figure 6 The logic shown in the constant tension open-loop control logic diagram establishes torque balance by increasing the output torque of variable motors 110 and 110' and fixed motors 136 and 136'. When the torque balance proximity switches 107 and 107' detect the balance signal, they energize the reversing valve XII 301 to release the brake. When the brake is already released (e.g., the drum is already in motion, or in the AOPS function mode where the initial speed is zero but the brake is already released), the system skips the above brake release step, and the speed and active compensation control module directly enters the active control mode. At this time, the total speed command value is generated by superimposing the operating handle signal and the attitude sensor (MRU) signal, and the encoders 137 and 137' detect the actual speed of the drum to form a speed closed-loop feedback. During active speed control, the control system continuously monitors the values ​​of the pressure sensors 212 and 204 in the two chambers of the active compensation motor 223, and determines whether the current active compensation output torque is within the set range based on the pressure difference. When the pressure difference is within the allowable range, the proportional servo valve 200 adjusts the active compensation motor 223 according to the speed control command; when the pressure difference exceeds the set threshold (e.g., ...), the control system adjusts the active compensation motor 223. Figure 9 As shown (taking 75% of the rated value as an example), the system maintains active compensation output and synchronously adjusts the swashplate angles of the variable motors 110 and 110' in the tension and passive compensation control modules to redistribute the overall load until the pressure difference returns to the allowable range, and then continues to execute speed control commands. Through this method, dynamic load distribution and coordinated control between the speed and active compensation control modules and the tension and passive compensation control modules are achieved. During the uniform speed operation phase of the drum, as the system acceleration α approaches zero, the required driving torque output of the active compensation motor 223 is reduced to primarily overcome the frictional torque Tf, thereby significantly reducing system energy consumption. Gearbox III 83, without a brake, is located on the other side of the drum 82.

[0047] The brake control module is responsible for controlling the actions of brakes 108 and 108'. The module includes brakes 108 and 108' located in the drum drive link. When the torque balance condition is met, it supplies or releases control oil pressure to the actuators of brakes 108 and 108' to raise or lower them. Before raising brakes 108 and 108', the brake control module outputs a pre-balancing torque through variable displacement motors 110 and 110' and fixed displacement motors 136 and 136'. After triggering torque balance proximity switches 107 and 107', the reversing valve XII 301 raises or lowers the brakes. The raising of brakes 108 and 108' is not direct, but is first achieved through… Figure 6 The logic shown in the constant tension open-loop control logic diagram increases the torque of variable motors 110 and 110' and fixed-displacement motors 136 and 136' until the torque balance proximity switches 107 and 107' detect a signal, indicating that the torque has reached the balance condition. Subsequently, the reversing valve XII 301 is energized, driving the brakes 108 and 108' to lift. The inlet pressure of the brake control module is limited by the pressure reducing valve III 300 to prevent it from exceeding the allowable brake pressure, and its pressure value is monitored by the pressure sensor VII 302. The pressure value determines whether the brake is lifted or closed.

[0048] The accumulator module plays a crucial role in energy storage, release, and providing emergency oil supply during wave compensation. For details of its structure and working principle, please refer to [link to relevant documentation]. Figure 3 (Schematic diagram of accumulator module) As shown, the core of this module is a piston accumulator 21, which is selectively connected to the system's high-pressure oil circuit via an isolation valve 35. During normal operation of the power station, the piston accumulator 21 stores energy and releases pressurized oil when needed by the system. When the power station malfunctions or fails, the piston accumulator 21 provides an emergency oil source to the system's pilot control circuit via the switching of the reversing valve V40. The pilot oil interface is shown below. Figure 3As shown in the diagram, this ensures the system retains basic control capabilities even under abnormal operating conditions. The piston accumulator 21 has a gas cylinder 25 connected in series on the gas side to improve gas-side stiffness and reduce oil-side pressure fluctuations. A magnetostrictive sensor 22 detects the initial piston position, pressure sensor I 24 monitors gas pressure in real time, and a safety valve 23 prevents excessive gas pressure. The oil-side isolation valve 35 is controlled by a reversing seat valve 38, two check valves 36 and 37 ensure stable control of the oil source pressure, and relief valve III 39 ensures the maximum pressure does not exceed the design value. To prevent excessive flow at the moment the isolation valve 35 opens, causing the piston to move too quickly, the system has a soft-start procedure: first, the reversing valve III 33 is energized, opening the reversing valve core I 31, and the flow is limited by the speed control valve 32; the isolation valve 35 can only be opened when the pressure signal difference between pressure sensor II 34 and pressure sensor III 50 is within 20 bar.

[0049] Under normal operating conditions of the power station, the emergency oil source is provided by the power pump station. The pressurized oil is sequentially controlled by one-way valve VIII 44 and one-way speed control valve 45, and then reduced to the rated pilot pressure required by the servo valve through pressure reducing valve I 46. The pressure of this pilot oil circuit is monitored in real time by pressure monitoring sensor 43, and relief valve IV 41 is used to provide overload protection for this oil circuit. When the power station malfunctions or fails, directional valve IV 40 is energized, switching the emergency oil source to be supplied to the system by piston accumulator 21. Among them, one-way valve IX 47 is used to prevent oil from flowing back into pressure reducing valve I 46 and causing unnecessary leakage, while one-way valves 138 and 138' are mainly used to flush and replenish oil to the high-pressure circuit through the low-pressure circuit after the power station is restarted.

[0050] The piston-type accumulator 21, when outputting control oil as an emergency oil source, limits the flow rate through a one-way speed control valve 45. When the power station module outputs oil, it works in conjunction with the pulsation absorption accumulators 53 and 53' to provide or absorb instantaneous large-flow oil sources and reduce pressure pulsations. Furthermore, the system includes multiple parallel accumulators 52, 52', and 52'' (expandable to multiple sets) to provide instantaneous large flow rates to the inlets of the fixed displacement pumps 1 and 1', and pulsation absorption accumulators 53 and 53' to reduce pressure pulsations at the pump outlet. The overflow valve VI 55 protects the high-pressure oil circuit pressure from exceeding the design pressure and directs the overflow liquid to the low-pressure circuit. The overflow valve V 54 protects the low-pressure circuit (dark blue solid line in the diagram) from exceeding the design pressure.

[0051] The power station module provides power to the system; its principle is detailed in [link to documentation]. Figure 2(Power station module schematic diagram). It primarily employs a redundant design, including fixed displacement pumps 1 and 1' driven by servo motors 2 and 2', used to supply oil to the system's high-pressure oil circuit. The system uses the signal from pressure sensor III 50 to control the pumps at constant pressure. Logic valves 5 and 5', in conjunction with directional valves 6 and 6', can isolate the outlet of a single pump in case of failure, preventing pressure differential from causing rapid drive damage to fixed displacement pump 1 or 1' even when servo motors 2 or 2' are de-energized. Relief valves 3 and 3' protect fixed displacement pumps 1 and 1' from exceeding their design pressure. Check valves 4 and 4' flush and replenish the system's high-pressure oil circuit. The low-pressure system is supplied by low-pressure pump 8 driven by motor 7; excess flow overflows through relief valve V 54 and is cooled by cooler 10. Check valve III 9 provides back pressure to fixed displacement pumps 1 and 1' to prevent their housings from venting. Check valve IV 11 ensures that the system pressure relief circuit flows unidirectionally to oil tank 13, while check valve V 12 provides replenishment oil when the low-pressure line pressure is too low (such as in emergency release conditions).

[0052] In summary, the system of this invention achieves highly efficient active-passive wave compensation through the meticulous design and coordinated control of the aforementioned modules. The tension and passive compensation control module bears the main static load and works in conjunction with the accumulator module to absorb and release energy, significantly reducing the system's installed power. The speed and active compensation control module responds quickly to dynamic demands, providing precise speed control. Under the coordination of the control system, both modules coordinate the control of static and dynamic compensation torques based on the detection signals from the encoders 137 and 137' of the drum 82 and the force load sensor 400 of the wire rope, jointly ensuring the stability, safety, and efficiency of deep-sea hoisting operations under complex sea conditions.

[0053] In conclusion, the above description is only a preferred embodiment of the present invention and is not intended to limit the present invention. Any modifications, equivalent substitutions, improvements, etc., made within the spirit and principles of the present invention should be included within the protection scope of the present invention.

Claims

1. An active-passive wave compensation hydraulic control system for deep-sea lifting, characterized in that, The system includes: The speed and active compensation control module includes an active compensation motor (223), whose inlet is connected to the high-pressure oil circuit of the system via a proportional servo valve (200), and its inlet pressure is stabilized via an oil source selection passage and a pressure compensation passage. The output of the active compensation motor (223) drives the drum (82) via gearbox III (83) to provide dynamic compensation torque. The tension and passive compensation control module includes a series-connected variable displacement motor (110, 110') and a fixed displacement motor (136, 136'), the inlet of which is connected to the high-pressure oil circuit of the system and the pilot pressure is kept stable through the pilot stabilization passage; the output of the tension and passive compensation control module drives the drum (82) through the gearbox (80, 80') to output static load compensation torque; The accumulator module includes a piston accumulator (21) that is selectively connected to the high-pressure oil circuit of the system via an isolation valve (35), and is switched via a reversing valve IV (40) when the power station fails, so that the piston accumulator (21) provides an emergency oil source to the system pilot control circuit; The power station module includes a fixed displacement pump (1, 1') driven by a servo motor (2, 2') to supply oil to the system's high-pressure oil circuit; The tension and passive compensation control module and the speed and active compensation control module are both mechanically connected to the drum (82). The encoder (137, 137') is set at the shaft end. The shaft end belongs to the fixed motor (136, 136') or variable motor (110, 110') connected to the gear transmission of the drum (82). The encoder is used to characterize the rotational speed of the drum (82) and the detection signal of the force load sensor (400) of the wire rope set on the drum (82) to coordinate and control the static load compensation torque and dynamic compensation torque output by the two modules.

2. The system according to claim 1, characterized in that, The swashplate angle of the variable motors (110, 110') is controlled by a closed-loop position control system consisting of servo valves (123, 123') and swashplate angle sensors (111, 111') to achieve precise adjustment of the static load compensation torque.

3. The system according to claim 1, characterized in that, The variable displacement motors (110, 110') and fixed displacement motors (136, 136') are connected in series with a displacement ratio of 1:1, thereby forming a static load compensation unit capable of outputting torque from zero to maximum.

4. The system according to claim 1, characterized in that, The high-pressure inlet of the tension and passive compensation control module is controlled to open / close the pressure circuit through logic valves (101, 101', 131, 131'). The opening or closing action of the logic valves (101, 101', 131, 131') controls whether the high-pressure inlet is connected to the high-pressure oil circuit of the system.

5. The system according to claim 1, characterized in that, The electromagnetic reversing valve V (214) of the speed and active compensation control module can switch the active compensation motor (223) to active control mode when energized and switch it to low differential pressure floating mode when de-energized, so as to adapt to the operation requirements of constant tension control or active speed control, and the pilot circuit flushing and oil replenishment is provided by the one-way valve XXIX (222).

6. The system according to claim 1, characterized in that, The speed and active compensation control module enables the active compensation motor (223) to automatically select the higher pressure oil source from the two parallel one-way valves (220, 221) as the oil supply source through a parallel unidirectional oil source selection path.

7. The system according to claim 1, characterized in that, The speed and active compensation control module includes a pressure compensator composed of shuttle valve I (201) and pressure reducing valve II (202) for pressure compensation of the oil supply to the active compensation motor (223). A two-way balance valve (203, 210) is provided on the oil return side of the active compensation motor (223) to adjust its motion state during operation, thereby ensuring the controllability and stability of the active compensation motor (223).

8. The system according to claim 1, characterized in that, It also includes a brake control module, which includes brakes (108, 108') disposed in the drum drive link, for raising or braking the brakes (108, 108') by supplying or releasing control oil pressure to the actuator of the brakes (108, 108') when the torque balance condition is met.

9. The system according to claim 8, characterized in that, Before the brake (108, 108') is lifted, the brake control module outputs a pre-balance torque through the variable motor (110, 110') and the fixed motor (136, 136'). After triggering the torque balance proximity switch (107, 107'), the reversing valve XII (301) realizes the lifting or braking of the brake.

10. The system according to claim 1, characterized in that, When the piston accumulator (21) outputs control oil as an emergency oil source, it limits the flow rate through a one-way speed regulating valve (45). When the power station module outputs oil, it works in conjunction with the pulsation absorption accumulator (53, 53') to provide or absorb instantaneous large flow oil sources and reduce pressure pulsation.

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