An electro-hydraulic mixed drive slewing system multi-way valve system and control method of engineering machinery

Abstract

The application discloses an engineering machinery electro-hydraulic mixed drive rotating system multi-way valve system and a control method. The system comprises a valve body, EA, A1, B1, P1, PG and T1 oil channel ports; a control valve assembly comprises a main reversing valve, a first reversing valve and a second reversing valve; the main reversing valve has Z1, Z2, Z3, Z4, Z5 and Z6 ends; the first reversing valve has C1, C2 and C3 ends; the second reversing valve has D1, D2 and D3 ends; the C1 end is connected with the Z1 end through a first hydraulic oil path; the D1 end is connected with the Z6 end through a second hydraulic oil path; the PG oil channel port is connected with the C2 end through a first connecting oil path; the PG oil channel port is connected with the D2 end through a second connecting oil path; the EA oil channel port is connected with the Z3 end through a third connecting oil path; the A1 oil channel port is connected with the Z2 end through a fourth connecting oil path; and the B1 oil channel port is connected with the Z4 end through a fifth connecting oil path.

Description

Technical Field

[0001] This application belongs to the technical field, specifically relating to a multi-way valve system and control method for an electro-hydraulic hybrid drive rotary system of engineering machinery. Background Technology

[0002] As the core actuator of the entire machine operation, the slewing mechanism of construction machinery undertakes key operations such as hoisting, excavation, leveling, and turning. Its smoothness of operation, control precision, energy consumption level and reliability directly determine the overall operating performance of the construction machinery.

[0003] First, traditional slewing systems mostly employ a single hydraulic or electric braking mode, lacking a dual braking redundancy protection mechanism. During high-speed, heavy-load slewing deceleration, if a single braking mechanism fails, the backup braking scheme cannot be quickly triggered, easily leading to major safety hazards such as slewing overspeed and runaway. This makes them unsuitable for complex and high-risk operating scenarios in construction machinery, such as high-altitude, heavy-load, and off-center loading, resulting in low equipment operational fault tolerance and safety. Furthermore, existing slewing control systems have low accuracy in identifying operating conditions, failing to achieve precise adaptation across all operating conditions. Traditional control schemes often rely on a single speed signal as the basis for determining operating conditions, without combining load torque, hydraulic flow, and other multi-dimensional parameters for comprehensive judgment, making it difficult to accurately distinguish typical operating conditions such as slewing start-up, acceleration, constant speed, and deceleration. During the slewing start-up phase, the inability to anticipate heavy-load starting conditions and the lack of a pre-energy storage buffer mechanism easily lead to problems such as start-up jerking, shaking, and insufficient power. During the acceleration phase, unreasonable power output matching easily results in response lag and low acceleration efficiency. During the constant-speed operation phase, the inability to achieve optimal switching of the power system generally results in continuous overflow and energy consumption of the hydraulic system. During the deceleration phase, relying solely on single hydraulic throttling braking or motor braking cannot match the characteristics of large rotational inertia loads, easily causing faults such as large braking impact, stationary deviation, and slippage, seriously affecting the precision of fine-tuning operations. Secondly, existing electro-hydraulic hybrid drive slewing control modes mostly adopt a fixed power output method, which cannot dynamically switch between single and dual drive modes according to working conditions. During stable, uniform rotation, the hydraulic system and motor operate synchronously, resulting in prolonged high-pressure overflow of the hydraulic pump, generating significant amounts of wasted energy. This also causes rapid increases in hydraulic oil temperature, accelerated aging of seals, and increased wear on pipelines. In the deceleration phase, where inertial energy is most concentrated, current technologies often employ pure hydraulic throttling for heat dissipation, failing to achieve dual-path energy recovery of both electrical and hydraulic energy. A large amount of inertial mechanical energy is wasted, leading to persistently high overall equipment energy consumption and hindering the energy-saving advantages of electro-hydraulic hybrid drive systems. Furthermore, during deceleration, a single braking mechanism bears all the impact loads. Prolonged alternating loads can easily cause fatigue damage to core components such as the slewing bearings, reducer, and valve assembly, significantly shortening the equipment's lifespan. Summary of the Invention

[0004] This application provides a multi-way valve system and control method for an electro-hydraulic hybrid drive slewing system for engineering machinery, addressing the aforementioned technical problems. Traditional slewing systems mostly employ a single hydraulic braking or single electric braking mode, lacking a dual braking redundancy protection mechanism. During high-speed, heavy-load slewing deceleration, if a single braking mechanism malfunctions, the backup braking scheme cannot be quickly triggered, easily leading to significant safety hazards such as slewing overspeed and runaway. Furthermore, existing electro-hydraulic hybrid drive slewing control modes mostly employ a fixed power output method, failing to dynamically switch between single and dual drive modes according to operating conditions.

[0005] The technical solution adopted in this application is as follows: A multi-way valve system for an electro-hydraulic hybrid drive rotary system of engineering machinery includes: The valve body has an EA oil port, an A1 oil port, a B1 oil port, a P1 oil port, a PG oil port, and a T1 oil port; the EA oil port is connected to the accumulator; the A1 oil port and the B1 oil port are respectively connected to the rotary mechanism; the P1 oil port, the T1 oil port, and the PG oil port are respectively connected to the pump. The control valve assembly includes a main directional valve, a first directional valve, and a second directional valve. The main directional valve has terminals Z1, Z2, Z3, Z4, Z5, and Z6. The first directional valve has terminals C1, C2, and C3. The second directional valve has terminals D1, D2, and D3. Terminals C1 and Z1 are connected via a first hydraulic circuit. Terminals D1 and Z6 are connected via a second hydraulic circuit. The PG port is connected to terminal C2 via a first connecting circuit. The PG port is connected to terminal D2 via a second connecting circuit. The EA port is connected to terminal Z3 via a third connecting circuit. The A1 port is connected to terminal Z2 via a fourth connecting circuit. The B1 port is connected to terminal Z4 via a fifth connecting circuit.

[0006] The multi-way valve system of the electro-hydraulic hybrid drive rotary system of engineering machinery of this application further includes a proportional switching valve, a check valve and a pressure reducing valve; the proportional switching valve has an E1 end and an E2 end; the pressure reducing valve has an F1 end and an F2 end; the E1 end and the Z3 end are connected through a sixth connecting oil circuit; the E2 end is connected through the check valve and the F2 end through a seventh connecting oil circuit; the F1 end is connected to the P1 oil passage through an eighth connecting oil circuit.

[0007] The multi-way valve system of the electro-hydraulic hybrid drive rotary system of engineering machinery of this application also includes a manual switching valve and a relief valve; the manual switching valve has a G1 end and a G2 end, the G1 end is connected to the EA oil passage through the ninth connecting oil circuit; the G2 end is connected to the T1 oil passage through the tenth connecting oil circuit; the relief valve has an H1 end and an H2 end, the H1 end is connected to the EA oil passage through the eleventh connecting oil circuit, and the H2 end is connected to the T1 oil passage through the twelfth connecting oil circuit.

[0008] This application also relates to a multi-way valve control method for an electro-hydraulic hybrid drive rotary system of engineering machinery, based on the aforementioned multi-way valve system for an electro-hydraulic hybrid drive rotary system of engineering machinery, comprising: Real-time data collection of speed, load torque, and flow rate via sensors; The collected values ​​are filtered and noise reduced to remove instantaneous interference spikes. The pre-processed values ​​are compared with standard parameter values ​​to determine whether the working condition corresponding to the slewing operation is the starting condition, acceleration condition, constant speed condition or deceleration condition. The corresponding control strategy is matched according to different operating conditions, and the continuously adjustable multi-way valve switching and opening control command is generated according to the control strategy.

[0009] The specific steps for determining the operating condition corresponding to the slewing operation state include: Real-time speed data is obtained through a speed sensor, real-time load torque is obtained through a torque sensor, and real-time flow data is obtained through a flow sensor. When the speed is lower than a certain target value and the load torque increases significantly, it is determined to be a starting condition. At this time, the accumulator is charged with hydraulic oil and pressurized through the main pump. If the flow rate continues to increase before the target threshold is reached, it is determined to be an acceleration condition. At this time, the main pump maintains the maximum output displacement, and the motor and the motor drive the rotary mechanism at the same time. When both the speed and the load torque are in a stable phase, it is determined to be a constant speed operating condition. At this time, the motor is in low-voltage standby mode, and the rotary mechanism is driven by the motor alone. When the rotational speed begins to decrease and the load torque continues to increase, it is determined to be a deceleration condition. At this time, the motor and the rotary mechanism are driven simultaneously.

[0010] During startup, the pressure reducing valve is energized, connecting the F1 and F2 terminals. The hydraulic oil from the main pump outlet enters the valve body through the P1 oil passage, then flows through the F1 and F2 terminals of the pressure reducing valve into the check valve, and finally enters the EA oil passage to charge the accumulator with hydraulic oil.

[0011] During acceleration, the second directional control valve and the proportional switching valve are energized. Pilot hydraulic oil enters the valve body through the PG port, and then enters the Z6 port of the main directional control valve through the D2 and D1 ports of the second directional control valve. This causes the main directional control valve to switch positions, allowing the hydraulic oil in the accumulator to enter the valve body through the EA port. Then, it enters the E3 port of the main directional control valve through the E2 and E1 ports of the proportional switching valve. From the E2 port of the main directional control valve, it flows out into the A1 port, then into the motor, then into the B1 port, and finally into the Z4 port of the main directional control valve. Finally, it flows from the Z5 port of the main directional control valve back to the hydraulic oil tank through the T1 port.

[0012] Hydraulic oil in the accumulator enters the valve body through the EA oil passage. If the hydraulic oil pressure is higher than the set overflow pressure threshold of the overflow valve, the hydraulic oil enters the hydraulic oil tank connected to the T1 oil passage through the H2 end of the overflow valve. If the hydraulic oil pressure does not reach the set overflow pressure threshold of the relief valve, the hydraulic oil enters the E2 terminal of the proportional switching valve.

[0013] Under constant speed conditions, the second directional valve and the proportional switch valve are de-energized, and the hydraulic oil at the Z6 end of the main directional valve flows into the Z1 end of the second directional valve and flows to the oil tank from the Z3 end of the second directional valve. When the main directional valve switches to the neutral position, the hydraulic oil in the motor flows back to the oil tank through the Z5 end of the main directional valve. At this time, the slewing mechanism is driven by the motor alone.

[0014] During deceleration, the first directional valve and the proportional switching valve are energized. Pilot oil flows into the valve body through the PG oil passage, and then enters the Z1 end of the main directional valve through the C2 and C1 ends of the first directional valve, realizing the gear switching of the main directional valve. The oil enters the E2 and E1 ends of the proportional switching valve through the EA oil passage, and then enters the Z3 end of the main directional valve. It then enters the B1 oil passage through the Z4 end of the main directional valve and flows into the motor. Finally, it flows into the Z2 end of the main directional valve through the A1 oil passage and returns to the oil tank through the Z5 end of the main directional valve.

[0015] Due to the adoption of the above technical solution, the beneficial effects achieved by this application are as follows: 1. This application integrates multiple connecting valves within the valve body, enabling these valves to interact and achieve a hybrid mode where the motor and accelerator jointly drive the rotary mechanism. It replaces the control rotary valve structure with an independent electro-hydraulic hybrid drive rotary system multi-way valve unit, reducing pressure loss. Using the motor and accelerator as a hybrid power source, and the control unit as a medium, it reduces the number of times the rotary actuator is started, improving energy utilization. Furthermore, the EA oil passage in this application is connected to an accumulator and has a storage device, reducing the frequency of motor and pump starts. Secondly, this application also utilizes the manual switching valve described below to manually depressurize, ensuring the normal operation of the hydraulic system.

[0016] This application can automatically allocate the power ratio between electric motor drive and hydraulic pump drive based on different working conditions, namely the load, operating speed, and working requirements of the equipment's rotation operation. For light-load and fine-tuning conditions, electric motor drive is prioritized to avoid overflow losses in the hydraulic system; for heavy-load and high-speed conditions, the hydraulic pump provides the main power, with the electric motor providing auxiliary power for voltage stabilization. This precise power matching logic significantly reduces ineffective starts and frequent start-stop actions of the rotary actuator, avoiding power waste, mechanical shock, and energy waste caused by frequent actuator starts and stops. Through diversified power sources and intelligent control, on-demand energy supply is achieved, completely improving the problems of power redundancy and high energy consumption in traditional systems, significantly improving the overall energy utilization efficiency of the machine, and reducing equipment operating energy costs. Attached Figure Description

[0017] The accompanying drawings, which are included to provide a further understanding of this application and form part of this application, illustrate exemplary embodiments and are used to explain this application, but do not constitute an undue limitation of this application. In the drawings: Figure 1 This is a schematic diagram of the external structure of a multi-way valve control system for an electro-hydraulic hybrid drive rotary system of engineering machinery, according to one embodiment of this application. Figure 2 This is a schematic diagram of the structure of a multi-way valve control system for an electro-hydraulic hybrid drive rotary system of engineering machinery, according to one embodiment of this application; Figure 3 This is a flow path diagram of hydraulic oil under the starting condition of a multi-way valve controlled electro-hydraulic hybrid drive rotary system for engineering machinery, according to one embodiment of this application. Figure 4 This is a flow path diagram of hydraulic oil under acceleration conditions controlled by a multi-way valve in an electro-hydraulic hybrid drive rotary system for engineering machinery, according to one embodiment of this application. Figure 5 This is a flow path diagram of hydraulic oil under uniform speed operation under multi-way valve control in an electro-hydraulic hybrid drive rotary system of engineering machinery according to one embodiment of this application. Figure 6 This is a flow path diagram of hydraulic oil under deceleration conditions controlled by a multi-way valve in an electro-hydraulic hybrid drive rotary system for engineering machinery, according to one embodiment of this application. In the picture, 1. Main directional valve; 2. Proportional switching valve; 3. First directional valve; 4. Check valve; 5. Pressure reducing valve; 6. Manual switching valve; 7. Second directional valve; 8. Relief valve; 9. Valve body. Detailed Implementation

[0018] To more clearly illustrate the overall concept of this application, a detailed explanation is provided below with reference to the accompanying drawings.

[0019] Many specific details are set forth in the following description in order to provide a full understanding of this application. However, this application may also be implemented in other ways different from those described herein. Therefore, the scope of protection of this application is not limited to the specific embodiments disclosed below.

[0020] Furthermore, it should be understood in the description of this application that the terms "top", "bottom", "inner", "outer", "axial", "radial", "circumferential", etc., indicate the orientation or positional relationship based on the orientation or positional relationship shown in the accompanying drawings, and are only for the convenience of describing this application and simplifying the description, and are not intended to indicate or imply that the device or element referred to must have a specific orientation, or be constructed and operated in a specific orientation, and therefore should not be construed as a limitation of the present invention.

[0021] In this application, unless otherwise expressly specified and limited, the terms "installation," "connection," "linking," and "fixing," etc., should be interpreted broadly. For example, they can refer to a fixed connection, a detachable connection, or an integral part; they can refer to a mechanical connection, an electrical connection, or a communication connection; they can refer to a direct connection or an indirect connection through an intermediate medium; they can refer to the internal communication of two components or the interaction between two components. Those skilled in the art can understand the specific meaning of the above terms in this invention according to the specific circumstances.

[0022] In this application, unless otherwise expressly specified and limited, the "above" or "below" of the second feature can mean that the first and second features are in direct contact, or that the first and second features are in indirect contact through an intermediate medium. In the description of this specification, references to terms such as "implementation," "example," "a particular embodiment," "example," or "specific example," etc., indicate that the specific feature, structure, material, or characteristic described in connection with that embodiment or example is included in at least one embodiment or example of this application. In this specification, the illustrative expressions of the above terms do not necessarily refer to the same embodiment or example. Furthermore, the specific features, structures, materials, or characteristics described can be combined in any suitable manner in one or more embodiments or examples.

[0023] Example 1 A multi-way valve system for an electro-hydraulic hybrid drive rotary system of engineering machinery, such as Figure 1-6 As shown, it includes: Valve body 9 has EA oil port, A1 oil port, B1 oil port, P1 oil port, PG oil port and T1 oil port; EA oil port is connected to the accumulator; A1 oil port and B1 oil port are respectively connected to the rotary mechanism; P1 oil port, T1 oil port and PG oil port are respectively connected to the pump. The control valve assembly includes a main directional valve 1, a first directional valve 3, and a second directional valve 7. The main directional valve 1 has Z1, Z2, Z3, Z4, Z5, and Z6 terminals. The first directional valve 3 has C1, C2, and C3 terminals. The second directional valve 7 has D1, D2, and D3 terminals. Terminals C1 and Z1 are connected by a first hydraulic circuit. Terminals D1 and Z6 are connected by a second hydraulic circuit. The PG oil passage is connected to terminal C2 by a first connecting oil circuit. The PG oil passage is connected to terminal D2 by a second connecting oil circuit. The EA oil passage is connected to terminal Z3 by a third connecting oil circuit. The A1 oil passage is connected to terminal Z2 by a fourth connecting oil circuit. The B1 oil passage is connected to terminal Z4 by a fifth connecting oil circuit.

[0024] like Figure 1 As shown, the left end face of the valve body 9 is connected to the first reversing valve 3 and the second reversing valve 7, as well as the check valve 4 and the EA oil passage. The first reversing valve 3, the second reversing valve 7, the check valve 4 and the oil passage EA are respectively connected to the oil passage inside the valve body 9. The right end face of the valve body 9 is connected to the A1 oil passage and the B1 oil passage. The A1 oil passage and the B1 oil passage are respectively connected to the oil passage inside the valve body 9. The upper end face of valve body 9 is connected to pressure test connector PA, main directional valve 1 and proportional switch valve 2. Pressure test connector PA, main directional valve 1 and proportional switch valve 2 are respectively connected to the oil passage inside valve body 9. The rear end face of valve body 9 is provided with oil passage T1, oil passage P1 and oil passage PG. T1, P1 and PG oil passages are respectively connected to the oil passage inside valve body 9. One side output shaft of motor is connected to rotary reducer, and one side input shaft of motor is connected to motor. Motor is connected to power battery through bidirectional DC-DC inverter. The motor and motor can jointly drive the rotary mechanism to rotate. This application integrates multiple connecting valves within the valve body 9, enabling these valves to interact and achieve a hybrid mode where the motor and accelerator jointly drive the rotary mechanism. It replaces the control rotary valve structure with an independent electro-hydraulic hybrid drive rotary system multi-way valve unit, reducing pressure loss. Using the motor and accelerator as a hybrid power source, and the control unit as a medium, it reduces the number of times the rotary actuator is started, improving energy utilization. Furthermore, the EA oil passage in this application is connected to an accumulator and has a storage device, reducing the frequency of motor and pump starts. Secondly, this application also uses the manual on / off valve 6 described below to manually relieve pressure, ensuring the normal operation of the hydraulic system.

[0025] In this application, oil port A1 is connected to motor port A, oil port B1 is connected to motor port B, oil port T1 is connected to the bypass return port of the pump, oil port P1 is connected to the main pump outlet of the pump, and hydraulic oil enters the valve body from the pump through oil port P1. Oil port PG is connected to the pilot port of the pump. Moreover, oil port EA is connected to the inlet ports of relief valve 8, manual switch valve 6 and proportional switch valve 2 through oil passages in valve body 9. That is, oil port EA is connected to the H1 end of relief valve 8 through the first connecting oil passage in valve body 9, oil port EA is also connected to the G1 of manual switch valve 6 through the ninth connecting oil passage, oil port EA is also connected to the E2 of proportional switch valve 2 through the third connecting oil passage, and then connected to the Z3 of main directional valve 1 through proportional switch valve 2. The other end of EA is also connected to an accumulator. The oil outlet of the overflow valve 8 and the oil outlet of the manual switch valve 6 are respectively connected to T1. That is, the H2 end of the overflow valve 8 is connected to the oil port of T1 through the twelfth connecting oil circuit, and the G2 of the manual switch valve 6 is connected to the oil port of T1 through the tenth connecting oil circuit.

[0026] The main directional valve 1 has six ports: Z1, Z2, Z3, Z4, Z5, and Z6. Z1 is connected to C1 of the first directional valve 3 via a first oil circuit; Z2 is connected to A1 via a fourth connecting oil circuit; Z3 is connected to E1 of the proportional switch valve 2 via a sixth connecting oil circuit; Z4 is connected to B1 via a fifth connecting oil circuit; Z5 is connected to T1 via a first pipeline; and Z6 is connected to D1 of the second directional valve 7 via a second oil circuit.

[0027] The first directional valve 3 has three ports, namely C1, C2 and C3; C1 is connected to Z1 through the first oil passage, C2 is connected to PG through the first connecting oil passage, and C3 is connected to T1 through the second pipeline.

[0028] The second directional valve 7 has three ports: D1, D2 and D3. D1 is connected to Z6 of the main directional valve 1 through the second oil circuit, D2 is connected to the PG oil passage through the second connecting oil circuit, and D3 is connected to the T1 oil passage through the third pipeline.

[0029] The multi-way valve system of the electro-hydraulic hybrid drive rotary system of engineering machinery of this application also includes a proportional switching valve 2, a check valve 4 and a pressure reducing valve 5; the proportional switching valve 2 has an E1 end and an E2 end; the pressure reducing valve 5 has an F1 end and an F2 end; the E1 end and the Z3 end are connected through a sixth connecting oil circuit; the E2 end is connected through the check valve 4 and the F2 end is connected through a seventh connecting oil circuit; the F1 end is connected to the P1 oil passage through an eighth connecting oil circuit.

[0030] The proportional switching valve 2 has two ports, namely E1 and E2. The E1 port is connected to the Z3 port of the main directional valve 1 through the sixth connecting oil circuit, and the E2 port is connected to the oil outlet of the check valve 4 through the seventh connecting oil circuit.

[0031] The pressure reducing valve 5 has two ports, namely F1 and F2; F1 is connected to P1 through the eighth connecting oil circuit, and F2 is connected to the oil inlet of the check valve 4 through the seventh connecting oil circuit.

[0032] The multi-way valve system of the electro-hydraulic hybrid drive rotary system of the engineering machinery of this application also includes a manual switch valve 6 and an overflow valve 8; the manual switch valve 6 has a G1 end and a G2 end, the G1 end is connected to the EA oil passage through the ninth connecting oil circuit; the G2 end is connected to the T1 oil passage through the tenth connecting oil circuit; the overflow valve 8 has an H1 end and an H2 end, the H1 end is connected to the EA oil passage through the eleventh connecting oil circuit, and the H2 end is connected to the T1 oil passage through the twelfth connecting oil circuit.

[0033] The manual switch valve 6 has two ports, namely G1 and G2. G1 is connected to the third connecting pipe connected to the EA oil passage through the ninth connecting oil passage, and G2 is connected to the T1 oil passage through the tenth connecting oil passage.

[0034] The overflow valve 8 is connected to the connecting pipeline that is connected to the EA oil passage. One end of the connecting pipeline is connected to the third connecting oil passage, and the other end is connected to the T1 oil passage.

[0035] Traditional rotary hydraulic systems often employ an independently controlled rotary valve combination structure, resulting in complex pipeline layouts, numerous valve group throttling channels, and cumbersome oil circuit transition nodes. During the transmission and regulation of hydraulic oil, throttling losses, pressure losses along the flow path, and local eddy current losses are easily generated, leading to a severe decrease in the effective working pressure of the system. This not only reduces the efficiency of hydraulic transmission but also causes problems such as hydraulic oil overheating and excessive system temperature rise.

[0036] This application replaces the original decentralized rotary valve control structure with an independent electro-hydraulic hybrid drive rotary system multi-way valve unit. This highly integrates the functions of pressure regulation, reversing, throttling, and pressure stabilization in rotary control, simplifying the system's hydraulic circuit structure and reducing redundant structures in pipeline connections and valve group throttling. The hydraulic oil flow path is shorter and the flow channel is smoother, significantly reducing throttling and turbulence losses within the valve group, effectively reducing overall system pressure loss, ensuring efficient conversion of hydraulic system input pressure into the rotary mechanism's operating power, and avoiding ineffective pressure loss. Simultaneously, stable pressure output reduces system pressure fluctuations, lowers hydraulic shock, improves the smoothness of rotary operation, and structurally enhances the hydraulic system's transmission efficiency, reducing energy waste under no-load and light-load conditions. This application can automatically allocate the power ratio between electric motor drive and hydraulic pump drive based on different working conditions, namely the load, operating speed, and working requirements of the equipment's rotation operation. For light-load and fine-tuning conditions, electric motor drive is prioritized to avoid overflow losses in the hydraulic system; for heavy-load and high-speed conditions, the hydraulic pump provides the main power, with the electric motor providing auxiliary power for voltage stabilization. This precise power matching logic significantly reduces ineffective starts and frequent start-stop actions of the rotary actuator, avoiding power waste, mechanical shock, and energy waste caused by frequent actuator starts and stops. Through diversified power sources and intelligent control, on-demand energy supply is achieved, completely improving the problems of power redundancy and high energy consumption in traditional systems, significantly improving the overall energy utilization efficiency of the machine, and reducing equipment operating energy costs.

[0037] Furthermore, as the core power components of the rotary system, the frequent start-stop of the motor and hydraulic pump is a major factor leading to accelerated wear, increased failure rate, and shortened service life. Traditional rotary systems lack energy storage and buffer structures; each rotary operation requires the motor and pump to directly start and perform work for starting, stabilizing, and replenishing pressure. High-frequency start-stop causes current surges in the motor, repeated pressure on the pump's volumetric cavity, and frequent stress aging of the sealing structure. Long-term operation can easily lead to faults such as motor overheating, pump wear, and increased internal leakage, significantly increasing equipment maintenance costs.

[0038] This application incorporates a dedicated energy storage device that can store excess hydraulic and electrical potential energy under no-load and light-load conditions. During equipment start-up, sudden load changes, and pressure replenishment, the energy storage device rapidly releases energy to assist the system, eliminating the need for frequent motor and hydraulic pump restarts. On one hand, this significantly reduces the start-stop frequency and operating condition switching frequency of motors and hydraulic pumps, effectively avoiding damage to core components from current surges and alternating mechanical loads, delaying component aging and wear, and significantly extending the service life of core components such as motors, hydraulic pumps, and rotary actuators, thereby reducing equipment failure rates and subsequent maintenance costs. On the other hand, the energy storage device can compensate for system pressure fluctuations in real time, stabilizing system operating pressure, avoiding sudden changes in speed and pressure caused by power start-stop, improving the uniformity and stability of rotary operations, and adapting to high-precision and high-stability operational requirements.

[0039] In addition, this application adds a manual on / off valve 6, which, in addition to automatic pressure relief, also enables manual pressure release. This achieves a dual safety protection mechanism. During normal equipment operation, the safety valve can monitor the system pressure in real time and, in conjunction with the automatic pressure relief structure, automatically unload overpressure, stabilizing the system pressure within the safe threshold. In special operating conditions such as equipment shutdown for maintenance, fault diagnosis, or failure of automatic pressure relief, residual system pressure can be quickly released manually, eliminating the potential for pressure buildup in pipelines and valve groups, and providing safe operating conditions for equipment maintenance and fault handling.

[0040] Example 2 This application also relates to a multi-way valve control method for an electro-hydraulic hybrid drive rotary system of engineering machinery, based on the aforementioned multi-way valve system for an electro-hydraulic hybrid drive rotary system of engineering machinery, comprising: Step S1: Real-time data collection of rotational speed, load torque, and flow rate using sensors; The sensors can be a speed sensor to collect real-time speed, a torque sensor to collect load torque, and a flow sensor to collect hydraulic oil flow.

[0041] Step S2: Filter and reduce noise based on the collected values ​​to remove instantaneous interference spikes, and compare the pre-processed values ​​with the standard parameter values ​​to determine whether the working condition corresponding to the slewing operation is the starting condition, acceleration condition, constant speed condition or deceleration condition. Step S3: Match the corresponding control strategy according to different operating conditions, and generate continuously adjustable multi-way valve switching and opening control commands according to the control strategy.

[0042] This application uses sensors to collect multi-dimensional operating parameters such as rotational speed, load torque, and flow rate in real time. Combined with filtering and noise reduction processing to eliminate instantaneous interference spikes, it effectively avoids misjudgments caused by noise and sudden signals, ensuring the authenticity and validity of the data. Then, by comparing pre-processed parameters with standard parameters, it accurately distinguishes four typical slewing conditions: start-up, acceleration, constant speed, and deceleration. This achieves refined judgment of the operating status, adapting to the entire slewing operation process of construction machinery. Furthermore, it matches dedicated control logic to different slewing conditions, moving away from a single fixed control mode. The control scheme can be dynamically adjusted according to the operating conditions, conforming to the operating characteristics of the electro-hydraulic hybrid drive slewing system. This avoids adverse conditions such as instantaneous large impacts and sudden pressure changes, reducing the load impact on multi-way valves, hydraulic lines, slewing mechanisms, and motors, reducing component wear, and extending the service life of the entire machine and valve group. Based on real-time operating conditions, it dynamically adjusts the opening of the multi-way valves, rationally allocating hydraulic power and motor power, making the electro-hydraulic hybrid drive system more rationally matched and reducing energy loss. The operating conditions are divided into four types: start-up, acceleration, constant speed, and deceleration. Different control strategies are implemented under these four conditions. During the start-up phase, oil is added to the accumulator. During the acceleration phase, the rotary mechanism is driven by both the electric motor and the hydraulic motor. The motor outputs torque at maximum current, and the pump maintains maximum output displacement. The motor and pump serve as a hybrid power source, with the valve body 9 as the medium, reducing the number of times the rotary actuator is started. During the constant speed phase, the hydraulic system pump is in low-pressure standby mode and is driven solely by the motor. During the deceleration phase, the motor and pump work simultaneously to achieve efficient operation of the rotary mechanism, avoiding energy loss and instantaneous impact. The motor is responsible for precise and controllable electric braking, while the hydraulic system is responsible for large inertia damping and energy storage buffering, resulting in smooth combined force, sufficient torque, and no impact.

[0043] Further, determining the operating condition corresponding to the slewing operation involves the following specific steps: Real-time speed data is obtained through a speed sensor, real-time load torque is obtained through a torque sensor, and real-time flow data is obtained through a flow sensor. When the speed is lower than a certain target value and the load torque increases significantly, it is determined to be a starting condition. At this time, the accumulator is charged with hydraulic oil and pressurized through the main pump. If the flow rate continues to increase before the target threshold is reached, it is determined to be an acceleration condition. At this time, the main pump maintains the maximum output displacement, and the motor and the motor drive the rotary mechanism at the same time. When both the speed and the load torque are in a stable phase, it is determined to be a constant speed operating condition. At this time, the motor is in low-voltage standby mode, and the rotary mechanism is driven by the motor alone. When the rotational speed begins to decrease and the load torque continues to increase, it is determined to be a deceleration condition. At this time, the motor and the rotary mechanism are driven simultaneously.

[0044] In determining the starting condition, the system uses a dual-parameter linkage of low speed and sudden torque increase to accurately identify the initial operating state of the slewing mechanism as it starts from zero and overcomes static friction, avoiding the problem of misjudging the condition by relying solely on speed. Construction machinery experiences high static load resistance during slewing starts. This method accurately captures the starting condition and, in conjunction with the main pump pre-pressurizing and storing oil in the accumulator, pre-stores hydraulic power, effectively offsetting the high load resistance at startup. This completely solves the problems of start-up jerking, jamming, and insufficient power in traditional slewing systems, significantly improving operational smoothness.

[0045] In judging the acceleration condition, a dynamic parameter judgment logic is adopted, which continuously increases the speed and increases the flow rate simultaneously. This can capture the operator's intention to accelerate the operation in real time, accurately distinguish between constant speed idling and dynamic acceleration. During the acceleration phase, the main pump is controlled to maintain the maximum output displacement. Combined with the synchronous drive of the electric motor and hydraulic motor, the advantages of the hydraulic motor's high torque and high instantaneous power and the electric motor's high precision and fast response are fully utilized to achieve superimposed power output, quickly increase the slewing speed to the target threshold, significantly improve the efficiency of slewing operation, shorten the operation cycle, and improve the overall operation efficiency of the machine.

[0046] In the judgment of uniform speed operation, the load is stable and there is no instantaneous impact load during the uniform speed rotation stage. The control mode is switched to motor-only drive and hydraulic motor low-pressure standby, and the high power output of the hydraulic system is shut down to avoid overflow loss and mechanical friction loss of the hydraulic pump during continuous operation.

[0047] In determining the deceleration condition, the deceleration is not based on power output, but rather on the coordinated braking damping formed by the electric motor and hydraulic motor. The condition determination logic of decreasing speed and increasing torque can accurately capture the deceleration state due to rotational inertia. The hydraulic motor provides a large damping torque to counteract the large rotational inertia, while the electric motor provides high-precision electrical control torque to fine-tune the deceleration rate. This solves the pain points of traditional single hydraulic deceleration, such as large impact and insufficient deceleration torque of single motor, achieving linear, smooth, and slip-free deceleration throughout the entire rotational deceleration process. In the dual-drive coordinated working mode, the hydraulic motor can convert some of the inertial energy into hydraulic energy and store it in the accumulator, while the electric motor can convert the inertial energy into electrical energy to recharge the accumulator. This achieves dual-path energy recovery of hydraulic and electrical energy, avoiding the problem of traditional pure throttling deceleration where all energy is wasted as heat, and maximizing the overall energy utilization rate of the machine.

[0048] Furthermore, during startup, the pressure reducing valve 5 is energized, connecting the F1 and F2 ends of the pressure reducing valve 5. The hydraulic oil from the main pump outlet enters the valve body 9 through the P1 oil passage, then flows through the F1 and F2 ends of the pressure reducing valve 5 into the check valve 4, and then enters the EA oil passage to charge the accumulator with hydraulic oil.

[0049] The accumulator at the EA oil passage needs to be pre-charged with oil. When the pressure reducing valve 5 is energized and switched to the left position, the hydraulic oil from the oil pump flows into the valve body 9 through the P1 oil passage, then enters the F1 end of the pressure reducing valve 5 through the eighth connecting oil circuit, and then flows into the check valve 4 through the F2 end of the pressure reducing valve 5. Finally, it flows out of the EA oil passage through the outlet of the check valve 4 and into the accumulator. At the same time, the pressure sensor inside the accumulator at the PA oil passage is observed. When the pressure reaches 25MPa, the oil charging stops.

[0050] Furthermore, under acceleration conditions, the second directional valve 7 and the proportional switching valve 2 are energized, and the pilot hydraulic oil enters the valve body 9 through the PG oil passage. It then enters the Z6 end of the main directional valve 1 through the D2 and D1 ends of the second directional valve 7, causing the main directional valve 1 to switch gears. This allows the hydraulic oil in the accumulator to enter the valve body 9 through the EA oil passage, and then enters the E3 end of the main directional valve 1 through the E2 and E1 ends of the proportional switching valve 2. It then flows out from the E2 end of the main directional valve 1 into the A1 oil passage, then into the motor through the A1 oil passage, then into the B1 oil passage, and finally into the Z4 end of the main directional valve 1. Finally, it flows from the Z5 end of the main directional valve 1 to the T1 oil passage and returns to the hydraulic oil tank.

[0051] During the acceleration start-up phase, the second directional valve 7 and the proportional switching valve 2 are energized. The second directional valve 7 uses a two-position three-way solenoid switching mechanism. When the second directional valve 7 is switched to the right position, the proportional switching valve 2 is switched to the left position. Pilot oil flows into the valve body 9 through the PG oil passage, then flows into the D2 terminal of the second directional valve 7 through the second connecting oil circuit. It then flows into the Z6 terminal of the main directional valve 1 through the D1 terminal of the second directional valve 7 and the second oil circuit. The main directional valve 1 uses a three-position four-way directional valve. The valve core of the main directional valve 1 is pushed to move and switch to the left position. The oil in the accumulator enters the valve body 9 through the EA oil passage. At this time, the relief valve 8 plays a safety protection role. If the oil pressure is too high, the oil will overflow through the relief valve. The H2 end of the flow valve 8 enters the oil tank connected to the T1 oil passage through the twelfth connecting oil circuit. If the oil pressure does not reach the overflow pressure of the relief valve 8, the hydraulic oil enters the Z3 end of the main directional valve 1 through the proportional switch valve 2. The hydraulic oil flowing out from the Z2 end of the main directional valve 1 enters the A port of the hydraulic motor through the fourth connecting oil circuit through the A1 oil passage, driving the hydraulic motor to rotate. The hydraulic oil flows out from the B port of the hydraulic motor, enters the valve body 9 through the B1 oil passage, and flows into the Z4 end of the main directional valve 1 through the fifth connecting oil circuit. The hydraulic oil flowing out from the Z5 end of the main directional valve 1 flows through the first pipeline through the T1 oil passage and then flows into the oil tank. At this time, the motor and the rotary mechanism are driven to accelerate rotation.

[0052] This application configures the hydraulic oil in the accumulator to enter the valve body 9 through the EA oil passage. If the hydraulic oil pressure is higher than the set overflow pressure threshold of the relief valve 8, the hydraulic oil enters the hydraulic oil tank connected to the T1 oil passage through the H2 end of the relief valve 8. If the hydraulic oil pressure does not reach the set overflow pressure threshold of the relief valve 8, the hydraulic oil enters the E2 end of the proportional switching valve 2. By adopting a centralized oil supply path of accumulator hydraulic oil entering the valve through the EA oil passage, and relying on the two-stage diversion control logic of the pressure threshold judgment of the relief valve 8 and the on / off state of the proportional switching valve 2, the oil circuit working mode can be automatically switched according to the real-time hydraulic oil pressure. High pressure over-threshold automatic pressure relief and low pressure normal precise oil supply completely solve the problems of disordered oil supply pressure and single oil circuit control of traditional accumulators, realizes the adaptive and precise matching of hydraulic pressure, and adapts to the pressure requirements of the rotary system for start-up, stop, acceleration and deceleration under all working conditions.

[0053] Under constant speed conditions, the second directional valve 7 and the proportional switch valve 2 are de-energized, and the hydraulic oil at the Z6 end of the main directional valve 1 flows into the Z1 end of the second directional valve 7 and flows from the Z3 end of the second directional valve 7 to the oil tank. When the main directional valve 1 switches to the neutral position, the hydraulic oil in the motor flows back to the oil tank through the Z5 end of the main directional valve 1. At this time, the slewing mechanism is driven by the motor alone.

[0054] When the second directional valve 7 and the proportional switch valve 2 are de-energized and return to their left and right positions respectively, the hydraulic oil at the D6 end of the main directional valve 1 flows into the D1 end of the second directional valve 7. The hydraulic oil flowing out from the D3 end of the second directional valve 7 flows back to the oil tank through the T1 oil passage. Then, the valve core of the main directional valve 1 returns to its original position and switches to the neutral position. At this time, no hydraulic oil flows into the C3 end of the first directional valve 3, which uses a three-position four-way directional valve. The proportional switch valve 2, the check valve 44, and the pressure reducing valve 55 are interconnected. At this time, the rotary mechanism is driven by the motor to rotate rapidly. The A1 oil passage is connected to the Z2 end of the main directional valve 1 via the fourth connecting oil circuit. Hydraulic oil flows through the A1 oil passage to the Z2 end of the main directional valve 1, and then flows out from the Z5 end of the main directional valve 1 to the T1 oil passage, i.e., back to the oil tank. At this time, the hydraulic motor is not working, so the hydraulic oil flows from the B1 oil passage through the fifth connecting oil circuit to the Z4 end of the main directional valve 1, and then through the Z5 end of the main directional valve 1 to the T1 oil passage, i.e., back to the oil tank. The hydraulic oil at the Z6 end of the main directional valve 1 flows through the second oil circuit to the D1 end of the second directional valve 7, and then through the D3 end of the second directional valve 7 to the T1 oil passage, i.e., back to the oil tank.

[0055] During deceleration, the first directional valve 3 and the proportional switching valve 2 are energized. Pilot oil flows into the valve body 9 through the PG oil passage, and enters the Z1 end of the main directional valve 1 through the C2 and C1 ends of the first directional valve 3, realizing the gear switching of the main directional valve 1. The oil enters the E2 and E1 ends of the proportional switching valve 2 through the EA oil passage, and then enters the Z3 end of the main directional valve 1. It enters the B1 oil passage through the Z4 end of the main directional valve 1 and flows into the motor. Then it flows into the Z2 end of the main directional valve 1 through the A1 oil passage, and returns to the oil tank through the Z5 end of the main directional valve 1.

[0056] When the first directional valve 3 and the proportional switching valve 2 are energized, the first directional valve 3 is switched to the right position and the proportional switching valve 2 is switched to the left position. Pilot oil flows into the valve body 9 through the PG oil passage, then flows into the C2 terminal of the first directional valve 3 through the first connecting oil circuit. It then flows through the C1 terminal of the first directional valve 3 through the first oil circuit to the Z1 terminal of the main directional valve 1, pushing the valve core of the main directional valve 1 to move and switch to the right position. Oil in the accumulator enters the valve body 9 through the EA oil passage. At this time, the relief valve 8 acts as a safety protection mechanism. If the oil pressure is too high, the oil enters the oil tank connected to the T1 oil passage through the H2 terminal of the relief valve 8. If the oil pressure is too low... When the overflow pressure of the relief valve 8 is reached, the hydraulic oil flows from the EA oil passage through the third connecting oil circuit through the D2 and D1 terminals of the proportional switch valve 2 to the Z3 terminal of the main directional valve 1. The hydraulic oil flowing out from the Z4 terminal of the main directional valve 1 enters the B1 oil passage through the fifth connecting oil circuit and then enters the B port of the hydraulic motor, driving the hydraulic motor to rotate in the opposite direction. The oil flows out from the A port of the hydraulic motor, enters the valve body 9 through the A1 oil passage, flows into the Z2 port of the main directional valve 1 through the fourth connecting oil circuit, and flows into the oil tank through the first pipeline from the Z5 terminal of the main directional valve 1 through the T1 oil passage. At this time, the rotary actuator is driven to decelerate and rotate by the motor and the motor together.

[0057] For any parts not mentioned in this application, existing technologies may be used or referenced.

[0058] The various embodiments in this specification are described in a progressive manner. The same or similar parts between the various embodiments can be referred to each other. Each embodiment focuses on describing the differences from other embodiments.

[0059] The above description is merely an embodiment of this application and is not intended to limit the scope of this application. Various modifications and variations can be made to this application by those skilled in the art. Any modifications, equivalent substitutions, improvements, etc., made within the spirit and principles of this application should be included within the scope of the claims of this application.

Claims

1. A multi-way valve system for an electro-hydraulic hybrid drive rotary system of engineering machinery, characterized in that, include: The valve body (9) has an EA oil port, an A1 oil port, a B1 oil port, a P1 oil port, a PG oil port and a T1 oil port; the EA oil port is connected to the accumulator; the A1 oil port and the B1 oil port are respectively connected to the rotary mechanism; the P1 oil port, the T1 oil port and the PG oil port are respectively connected to the pump; The control valve assembly includes a main directional valve (1), a first directional valve (3), and a second directional valve (7); the main directional valve (1) has Z1, Z2, Z3, Z4, Z5, and Z6 ends; the first directional valve (3) has C1, C2, and C3 ends; the second directional valve (7) has D1, D2, and D3 ends; the C1 end and the Z1 end are connected by a first hydraulic circuit; the D1 end and the Z6 end are connected by a second hydraulic circuit; the PG oil passage is connected to the C2 end by a first connecting oil circuit; the PG oil passage is connected to the D2 end by a second connecting oil circuit; the EA oil passage is connected to the Z3 end by a third connecting oil circuit; the A1 oil passage is connected to the Z2 end by a fourth connecting oil circuit; and the B1 oil passage is connected to the Z4 end by a fifth connecting oil circuit.

2. The multi-way valve system of an electro-hydraulic hybrid drive rotary system for engineering machinery as described in claim 1, characterized in that, It also includes a proportional switching valve (2), a check valve (4) and a pressure reducing valve (5); the proportional switching valve (2) has an E1 end and an E2 end; the pressure reducing valve (5) has an F1 end and an F2 end; the E1 end is connected to the Z3 end through a sixth connecting oil circuit; the E2 end is connected to the check valve (4) and the F2 end is connected to the seventh connecting oil circuit; the F1 end is connected to the P1 oil passage through an eighth connecting oil circuit.

3. The multi-way valve system of the electro-hydraulic hybrid drive rotary system for engineering machinery as described in claim 1, characterized in that, It also includes a manual switch valve (6) and a relief valve (8); the manual switch valve (6) has a G1 end and a G2 end, the G1 end is connected to the EA oil passage through the ninth connecting oil circuit; the G2 end is connected to the T1 oil passage through the tenth connecting oil circuit; the relief valve (8) has an H1 end and an H2 end, the H1 end is connected to the EA oil passage through the eleventh connecting oil circuit, and the H2 end is connected to the T1 oil passage through the twelfth connecting oil circuit.

4. A multi-way valve control method for an electro-hydraulic hybrid drive rotary system of engineering machinery, based on the multi-way valve system of an electro-hydraulic hybrid drive rotary system of engineering machinery as described in any one of claims 1-3, characterized in that, include: Real-time data collection of speed, load torque, and flow rate via sensors; The collected values ​​are filtered and noise reduced to remove instantaneous interference spikes. The pre-processed values ​​are compared with standard parameter values ​​to determine whether the working condition corresponding to the slewing operation is the starting condition, acceleration condition, constant speed condition or deceleration condition. The corresponding control strategy is matched according to different operating conditions, and the continuously adjustable multi-way valve switching and opening control command is generated according to the control strategy.

5. The multi-way valve control method for an electro-hydraulic hybrid drive rotary system of engineering machinery as described in claim 4, characterized in that, The specific steps for determining the operating condition corresponding to the slewing operation state include: Real-time speed data is obtained through a speed sensor, real-time load torque is obtained through a torque sensor, and real-time flow data is obtained through a flow sensor. When the speed is lower than a certain target value and the load torque increases significantly, it is determined to be a starting condition. At this time, the accumulator is charged with hydraulic oil and pressurized through the main pump. If the flow rate continues to increase before the target threshold is reached, it is determined to be an acceleration condition. At this time, the main pump maintains the maximum output displacement, and the motor and the motor drive the rotary mechanism at the same time. When both the speed and the load torque are in a stable phase, it is determined to be a constant speed operating condition. At this time, the motor is in low-voltage standby mode, and the rotary mechanism is driven by the motor alone. When the rotational speed begins to decrease and the load torque continues to increase, it is determined to be a deceleration condition. At this time, the motor and the rotary mechanism are driven simultaneously.

6. The multi-way valve control method for an electro-hydraulic hybrid drive rotary system of engineering machinery as described in claim 5, characterized in that, During startup, the pressure reducing valve (5) is energized, so that the F1 end of the pressure reducing valve (5) is connected to the F2 end. The hydraulic oil from the main pump outlet enters the valve body (9) from the P1 oil passage, and then flows into the check valve (4) through the F1 and F2 ends of the pressure reducing valve (5), and then enters the EA oil passage to charge the accumulator with hydraulic oil.

7. The multi-way valve control method for an electro-hydraulic hybrid drive rotary system of engineering machinery as described in claim 5, characterized in that, Under acceleration conditions, the second directional valve (7) and the proportional switch valve (2) are energized. The pilot hydraulic oil enters the valve body (9) through the PG oil passage, and enters the main directional valve (1) through the D2 and D1 ends of the second directional valve (7). This causes the main directional valve (1) to switch gears, allowing the hydraulic oil in the accumulator to enter the valve body (9) through the EA oil passage. Then, it enters the main directional valve (1) through the E2 and E1 ends of the proportional switch valve (2). Then, it flows out from the E2 end of the main directional valve (1) and enters the A1 oil passage. Then, it enters the motor from the A1 oil passage, and then flows into the B1 oil passage and enters the Z4 end of the main directional valve (1). Finally, it flows from the Z5 end of the main directional valve (1) to the T1 oil passage and returns to the hydraulic oil tank.

8. The multi-way valve control method for an electro-hydraulic hybrid drive rotary system of engineering machinery as described in claim 7, characterized in that, The hydraulic oil in the accumulator enters the valve body (9) through the EA oil passage; If the hydraulic oil pressure is higher than the set overflow pressure threshold of the overflow valve (8), the hydraulic oil enters the hydraulic oil tank connected to the T1 oil passage through the H2 end of the overflow valve (8). If the hydraulic oil pressure does not reach the set overflow pressure threshold of the overflow valve (8), the hydraulic oil enters the E2 terminal of the proportional switching valve (2).

9. A multi-way valve control method for an electro-hydraulic hybrid drive rotary system of engineering machinery as described in claim 5, characterized in that, Under constant speed conditions, the second directional valve (7) and the proportional switch valve (2) are de-energized, and the hydraulic oil at the Z6 end of the main directional valve (1) flows into the Z1 end of the second directional valve (7) and flows from the Z3 end of the second directional valve (7) to the oil tank. When the main directional valve (1) switches to the neutral position, the hydraulic oil in the motor flows back to the oil tank through the Z5 end of the main directional valve (1). At this time, the slewing mechanism is driven by the motor alone.

10. The multi-way valve control method for an electro-hydraulic hybrid drive rotary system of engineering machinery as described in claim 1, characterized in that, Under deceleration conditions, the first directional valve (3) and the proportional switch valve (2) are energized. Pilot oil flows into the valve body (9) through the PG oil passage, and enters the Z1 end of the main directional valve (1) through the C2 and C1 ends of the first directional valve (3), thereby realizing the gear switching of the main directional valve (1). The oil enters the E2 and E1 ends of the proportional switch valve (2) through the EA oil passage and then enters the Z3 end of the main directional valve (1). It enters the B1 oil passage through the Z4 end of the main directional valve (1) and flows into the motor. Then it flows into the Z2 end of the main directional valve (1) through the A1 oil passage and returns to the oil tank through the Z5 end of the main directional valve (1).

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