© H.-H. Shen et al., Published by EDP Sciences, 2025

1 Preface

Direct-drive pump-controlled electro-hydraulic servo system is a hot research topic in recent years. Based on the advantages of high efficiency, environmental protection, high degree of integration and strong bearing capacity, it is widely used in national defense and industrial fields. In particular, the application in the control systems of industrial robots, missiles, high-precision lathes, construction machinery and other traditional control systems makes the research on direct-drive pump-controlled electro-hydraulic systems more in-depth [1]. It is pointed out in reference [2] that the pump-controlled servo system has the disadvantages of slow dynamic response and low synchronization accuracy [2]. It is pointed out in reference [3] that under the condition of low-speed pump control, the direct-drive electro-hydraulic system still has control difficulties such as nonlinear pump flow and large flow deviation, which makes it difficult for the variable-speed pump control system to track the actuator trajectory with high precision [3]. It is pointed out in reference [4] that that the pump control system has control difficulties such as high system dynamics order and strong nonlinearity in practical application [4]; It is pointed out in reference [5] that the direct-drive pump-controlled electro-hydraulic servo system has the disadvantages of slow dynamic response and low synchronization accuracy [5]. It is found that the root cause of the above problems is the unstable flow caused by the uncertainty of internal parameters or the disturbance of load in the control process. Therefore, it is of great significance to adapt a flow stability control method with strong anti-interference ability and low complexity to the direct-drive electro-hydraulic servo system to improve the system performance.

There are two methods to solve the stability of hydraulic flow at present. One is to use advanced control algorithm. Chen Guiquan of Sichuan University has designed a fuzzy controller, which adopts a dual-input and single-output fuzzy controller to adjust the output speed of the motor in time to achieve flow balance through the pressure error and the speed of error change [6]. Wu Mingchao of Yantai University designed an active disturbance rejection controller, which is a control model that expands the total disturbance into an additional state variable, so as to estimate and compensate it, so that the system has higher control accuracy when it encounters external disturbance [7]. Ding Yinting of Henan University of Science and Technology used a PID control algorithm based on wavelet neural network (WNN), which made the control system have higher tracking accuracy, stronger adaptability and better robustness [8]. These control algorithms can improve the intelligence level and adaptability of the whole system, but they usually have relatively complex structures, or rely heavily on high-precision mathematical modeling, and the parameter adjustment is complicated, and the construction and understanding of most algorithms require a high theoretical design foundation, so it is difficult to popularize them in practical applications [8]. The second method is to use components to control, which has the advantages of clear physical meaning, simple parameter adjustment and simple algorithm structure.

Among existing flow control valves, the core structure of proportional flow valves consists of a valve spool driven by a proportional electromagnet. The opening of the spool is controlled by electrical signals to regulate flow, which usually relies on precision electromagnetic drive components and feedback sensors. Structurally, they require supporting electronic control modules, with numerous external pipeline connections, resulting in high system complexity, and they have problems of dead zones and nonlinearity [9]. Pressure-compensated valves are mainly composed of a spool, spring, valve seat, etc. They maintain stable pressure difference across the valve port through the balance between spring force and pressure difference. Their structure is relatively simple but depends on the stiffness characteristics of the spring, making it difficult to adapt to variable load conditions. Moreover, they need to be used in conjunction with other valve components, with complex external pipeline connections, and are easily affected by installation layout [10]. Aiming at the flow instability problem of direct-drive pump-controlled electro-hydraulic systems in low-frequency and low-amplitude scenarios (such as precision assembly and slow feeding), this paper proposes a flow stabilizing valve with a special structure: through the symmetrical design of double annular grooves and the coordinated action of a ball valve and a spool valve, passive pressure compensation is achieved, enabling rapid flow balancing without complex electronic control. Compared with traditional valves, its innovations lie in: ① adopting a composite structure of “spool valve regulation + ball valve unloading” to balance stability and response speed; ② the axially equidistant design of double annular grooves to ensure the symmetry of pressure compensation; ③ integrating the oil passage system to reduce external pipelines and lower the risk of leakage. Through structural analysis, simulation, and experimental verification, the role of this valve in improving speed stability is clarified, providing a practical solution for low-frequency and low-amplitude working conditions [11–15].

2 The structure and working circuit analysis of flow stabilizing valve

As shown in Figure 1, the structure of the flow stabilizing valve mainly includes a valve cover, a valve body, a valve core, a ball valve, some retaining rings and various oil ports. Valve sleeve, slide valve core and ball valve core are installed in the valve body. Sealing ring is adopted between the valve sleeve and the valve body, and the spool of the slide valve can slide axially in the valve sleeve. There are a pair of inner annular grooves on the valve sleeve and a pair of outer annular grooves on the valve core, and the axial spacing between them should be strictly equal. The oil passage system A of the valve core is connected with the annular groove and the oil inlet, and the oil passage system B is connected with another annular groove and the valve cavity. The ball valve is installed on the valve seat formed at the top of the valve sleeve and pressed by the spring [16].

When a double-rod hydraulic cylinder is used to drive a load, its circuit is shown in Figure 2. The hydraulic pump needs to supply pressure oil from the right side to enter the 4-chamber and 3-chamber for oil return if the hydraulic cylinder moves from right to left to push the left load. In this case, oil is delivered from port 1 of bidirectional quantitative pump, and oil enters from port 5 of valve 1.The valve core is pushed to move to the left against the spring force inside the valve 1, and the channels 5 to 6 and 7 to 9 are opened. And then the oil flows through the 6 ports of the valve 1 to the 12 ports of the valve 2.When the pressure reaches a certain level, the oil overcomes the spring force to push the ball open, and then the oil enters the four cavities of the hydraulic cylinder through 13 ports to push the piston to the left. The hydraulic oil in the low pressure area 3 of the cylinder returns to the oil suction port 2 of the pump through the 7 ports and 9 ports of the valve 1 (7 and 9 have been communicated).

Assuming that there is a reason that the hydraulic oil flowing out of the high pressure port of the hydraulic cylinder increases the pressure of the low pressure port of the pump, then the increased pressure will move the spool in the flow stabilizing valve 1 to the left. A part of the oil in the pump flows out through the load unloading port and the load oil port of the flow stabilizing valve 1, so as to reduce the fluid resistance and then reduce the pressure in the load oil port of the flow stabilizing valve 2.The reduced pressure will make the spool of the flow stabilizing valve 2 move to the left, so as to achieve throttling through the load unloading port and reduce the pressure at the high pressure end of the pump. This pressure compensation can accurately control the release of high-pressure fluid in the hydraulic cylinder.

Fig. 1 Structure of flow stabilizing valve. 1. Valve cover 2. Ball valve spool 3. Valve sleeve 4. Seal ring 5. Annular groove 6. Valve core annular groove 7. Valve core 8. Retaining ring 9. Valve body 10. Oil-in 11. Core oil passage A 12. Oil source unloading port 13. Core oil passage B 14. Load unloading port 15. Liquid filling port 16. End cap 17. Load oil port.

Fig. 2 System circuit with flow stabilizing valve.

Fig. 3 Direct drive system simulation model of non-flow stabilizing valve. 1. Input signal 2. Motor 3. Pump 4. Safety valve 5. Safety valve 6. Hydraulic cylinder 7. Mass block 8. Disturbing force 9. Disturbance signal.

3 Simulation model construction and parameter design

According to the working principle of direct-drive pump-controlled electro-hydraulic servo system, in order to analyze the influence of load disturbance on the system, the system model with load disturbance is built in AMESim [17]. As shown in Figure 3, the input signal of the quantitative pump is a variable direction speed signal (1500 r/min). The system uses variable load signal to simulate the disturbance in practical application [18,19].

The simulation model of direct-drive pump-controlled electro-hydraulic servo system with flow stabilizing valve is shown in Figure 4. The simulation parameter settings are shown in Table 1.

Fig. 4 Direct drive electro hydraulic actuator simulation model with flow stabilizing valve. 1. Input signal 2. Motor 3. Pump 4. Check valve 5. Hydraulic cylinder 6. Check valve 7. Flow stabilizing valve 8. Flow stabilizing valve 9. Hydraulic cylinder 10. Mass block 11. Disturbing force 12. Disturbance signal.

4 The system simulation and result analysis

The input signal of direct drive system without flow stabilizing valve is shown in Figure 5. By analyzing the working principle of the flow stabilizing valve, it is known that the movement direction of the hydraulic cylinder of the system with the flow stabilizing valve is opposite to that of the system without the flow stabilizing valve. In order to compare the simulation results conveniently, the input signals of the two systems are equivalent and opposite.

In the simulation process, the pulse signal (Fig. 6a) is used to simulate that the system is suddenly impacted by the outside world. Step signal (Fig. 6b) is used to simulate the eccentric load of the hydraulic cylinder and the force changes. If the bias load changes continuously, the ramp signal (Fig. 6c) is used for simulation; The sinusoidal signal (Fig. 6d) is used to simulate the periodic change of load disturbance.

When the load is disturbed by the pulse signal, the displacement curve of the system without the flow stabilizing valve is shown in Figure 7a, and the velocity curve is shown in Figure 7b. Curve 1 represents the curve without disturbance, and curve 2 represents the curve with disturbance.

By analyzing the displacement curve shown in Figure 7a, it can be seen that when the load is not disturbed, it reaches the maximum displacement of 0.052 m for the first time at 1.28 s and returns to the initial position at 11.31 s. If the load is disturbed, it reaches the maximum displacement of 0.052 m for the first time at 2 s, and returns to the initial position at 12 s.There is a time difference of 0.72 s and 0.69 s. Through the analysis of the speed curve, it can be known that the speed changes when the load is disturbed and undisturbed during the movement of the piston of the hydraulic cylinder. Such as 0.5 s, the speed without disturbance is 0.034 m/s, and the speed with pulse disturbance is 0.015 m/s. The load disturbance causes the speed to decrease, so it can be seen that the load disturbance affects the speed stability in the direct drive system without flow stabilizing valve.

For the system with flow stabilizing valve, the displacement curve is shown in Figure 8a, and the velocity curve is shown in Figure 8b. By analyzing the displacement curve, it can be seen that whether there is load disturbance or not, it reaches 0.052 m at 1.4 s. As can be seen from the speed curve, the speed is basically stable. So, the speed stability of the system is improved after adding the flow stabilizing valve. But the system with flow stabilizing valve has a 3 s lag in the first hydraulic cylinder return process.

The system with flow stabilizing valve and the system without flow stabilizing valve are respectively added with step signal, ramp signal and sine signal to simulate various load disturbances, and their speed simulation curves are shown in Figures 9–11. Four conclusions can be drawn by analyzing and comparing the simulation curves.

(1) The speed difference of the system without flow stabilizing valve is 0.036 m/s at the 10th time when the system is disturbed by step signal. The speed difference of the system with flow stabilizing valve is only 0.0016 m/s at 13 s, but there is a lag of 3 s.

(2) The speed difference of the system without flow stabilizing valve is 0.021 m/s at the 20th when the system is disturbed by the slope signal. The speed difference of the system with flow stabilizing valve is 0.010 m/s at 22.8 s, but there is a lag of 2.8 s.

(3) Compared with the system without flow stabilizing valve, the coincidence degree of velocity curve of the system with flow stabilizing valve is much higher. The main reason is that the oil flowing from the high-pressure end of the hydraulic cylinder increases the pressure at the low-pressure end of the pump. The increased pressure causes the spool in the flow stabilizing valve 1 to move to the left, part of the fluid in the pump flows out through the load unloading port and the load oil port, The pressure in the load oil port of the flow stabilizing valve 2 is reduced, so that the valve core moves to the left. Using the load unloading port to realize throttling, and then reduce the pressure at the high pressure end of the pump. This pressure compensation can accurately control the release of high-pressure fluid in the hydraulic cylinder.

(4) The displacement curve and velocity curve are delayed for about 3 s in the system with flow stabilizing valve, The main reason is that after adding the flow stabilizing valve in the system, the oil flow path increases, which leads to the speed lag and the rapidity of system control decreases, but it does not affect the speed stability. The lag can be compensated for by using low-viscosity hydraulic oil and a feedforward control algorithm.

Fig. 5 Input signal of the system.

Fig. 6 Disturbance signal of load.

Fig. 7 Simulation curve of pulse disturbance signal with no flow stabilizing valve system.

Fig. 8 Simulation curve of pulse disturbance signal with flow stabilizing valve system.

Fig. 9 Velocity curve of step disturbance signal.

Fig. 10 Velocity curve of slope disturbance signal.

Fig. 11 Velocity curve of sine disturbance signal.

5 Experimental analysis of system dynamic characteristics

MS1-R series AC servo drive unit provided by Shenzhen Huichuan Technology Co., Ltd. was adopted in the experiment. Motor model: MS1H3-29C15CD-T331. Rated torque: 18.6 N m. Rated speed: 1500 r/min. Maximum speed: 3000 r/min. Drive model: SV630PT012. Power cable model: BMP-GA24. Communication cable model: BSC-CC24A.

The hydraulic pump adopts CMW-F206 gear motor provided by Hefei Changyuan Hydraulic Co., Ltd. instead of bidirectional quantitative gear pump, and its theoretical displacement is 6.4 ml/r. The oil replenishing valve adopts the cartridge valve provided by Ningbo Ketai Hydraulic Co., Ltd. and integrates the valve block.

Based on the position accuracy requirements of electro-hydraulic servo system, PCIE-1812 data acquisition card provided by Advantech Technology (China) Co., Ltd. is selected, and its performance indicators are as follows: analog input is differential 8-channel analog input; 16-bit AD precision, sampling frequency is 250 ks/s; The input impedance is 100 gΩ/350 pf; Sampling can be carried out by software, and the AD range is ±10 V, 5 V, ±2.5 V and ±1.25 V, respectively. A magnetostrictive displacement sensor provided by Shenzhen mirante Technology Co., Ltd. is selected as the linear displacement sensor, the model of which is THM-0450mm-P02-V010, with a magnetic ring type, the measuring range is 450 mm, the voltage is 24 V, and the voltage is 0–10 V DC. The resolution is up to 0.5 µm, the nonlinearity is 0.01%FS, the repeatability is 0.001%FS, and the maximum output load is 2 mA.

The hydraulic cylinder is provided by Dongguan Xiangli Hydraulic Components Manufacturing Co., Ltd., and its cylinder diameter is 63 mm; The diameter of piston rod is 45 mm; The maximum stroke is 250 mm. Guide rail adopts the linear guide rail of Shenzhen Reliance Automation Equipment Co., Ltd., and the model is LWL20BCS. Based on the above mechanism selection, combined with the rationality of the overall layout of the system, the system test platform as shown in Figures 12 and 13 is built, which shows the external structure of the system.

In order to analyze the dynamic characteristics of the system with flow stabilizing valve, sinusoidal signals of 0.1 Hz and 1 Hz, ramp signals of 20 mm/s and 30 mm/s, and step signals of 30 mm and 40 mm amplitude were respectively input into the test device, and that displacement signal is convert into a flow signal through the flow, and then converted into a given servo motor with a speed signal. In order to verify the performance of the flow stabilizing valve, the output voltage of the motor controller is changed at 0.8 seconds, which affects the speed of the motor to disturb the system. Through experiments, the response curve as shown in Figure 14 is obtained. Curve 1 is an input signal and curve 2 is a test response signal.

By analyzing the curve in Figure 14a and 14b, it can be seen that the response characteristics of the system are different with the frequency of the input sinusoidal signal. The tracking performance of the system is good when the sinusoidal signal is 0.1 Hz, and the tracking performance is obviously reduced when the sinusoidal signal is 1 Hz, and the stability and controllability are also reduced. The main reason is that the speed of the motor used in the test is too low, which affects the accuracy of the system [20].

By analyzing the curves in (c) and (d) of Figure 14, it can be seen that when the ramp signal with the speed of 20 mm/s is input, the system follows well, and when it is 30 mm/s, the error increases. This is mainly because with the increase of the speed of the system, the frequency of oil replenishment and oil discharge increases, but the oil replenishment system can't reach the required frequency, which leads to the pressure drop in the high-pressure chamber of the hydraulic cylinder, the speed drop of the piston rod and even the creeping situation, and the system error increases accordingly [21].

By analyzing the curve in (e) and (f) of Figure 14, it can be seen that the response speed of the system with a step amplitude of 30 mm is obviously better than that of the system with a step amplitude of 40 mm, and the steady-state error of the system becomes obviously larger when the step amplitude is 40 mm. It can be judged that the movement speed of the piston rod of the hydraulic cylinder is obviously reduced, and the main reason is that the oil replenishment system can not meet the demand.

It can be concluded from the above analysis that the dynamic characteristics of the direct-drive electro-hydraulic servo system with flow stabilizing valve meet the requirements of the system when the low-frequency signal is input, and the step response curve also verifies that the steady-state error of the system is less than 0.04 mm. However, with the increase of the amplitude and frequency of the input signal, the rapidity of the system decreases obviously and its steady-state error increases. Therefore, this system meets the requirements when the input signal is low frequency and low amplitude.

Fig. 12 Chart of Hydraulic power part.

Fig. 13 Chart of Electric control part.

Fig. 14 Experimental response curve.

6 Conclusion

In order to solve the problem of speed stability in the direct-drive pump-controlled electro-hydraulic servo system, this paper designs a flow stabilizing valve with a special structure, which can quickly adjust the flow into the hydraulic cylinder when the load in the direct-drive electro-hydraulic servo system is disturbed and improve the stability of the system. Through the structural design, working principle analysis, simulation analysis with or without flow stabilizing valve system and dynamic characteristic test analysis of the system with flow stabilizing valve, it can be seen that the speed of the system with flow stabilizing valve does not change obviously when the hydraulic cylinder is disturbed by the outside, and the speed stability is good, which can effectively eliminate the flow imbalance caused by load disturbance in the position control system and improve the position accuracy. But this system is only used when the input signal is low frequency and low amplitude. However, this system is only applicable to scenarios with low-frequency (0.1–0.5 Hz) and low-amplitude (≤30 mm) input signals, such as mold clamping of injection molding machines and feeding of precision assembly platforms. In high-frequency (>1 Hz) or high-amplitude (>30 mm) scenarios, such as joint driving of industrial robots and high-speed punch presses, its performance degrades significantly due to lag and oil replenishment limitations, thus requiring optimization in combination with active control algorithms.

Funding

This work was supported by the Guided Project Approval for Science and Technology Research in Henan Province (Grant No. 252102220079).

Conflicts of interest

The authors declare no conflicts of interest.

Data availability statement

The data that support the findings of this study are available from the corresponding author (Huanhuan Shen, E-mail: shenhuanhuan@zzuit.edu.cn) upon reasonable request. For data involving the National Bureau of Statistics, permission from the National Bureau of Statistics must be obtained simultaneously.

Author contribution statement

All authors' contributions are clear: Huanhuan Shen is responsible for research conceptualization, writing of the paper's original draft, funding application, and project management; Ruolan Liu is responsible for experiment implementation and data analysis; Geqiang Li is responsible for paper review and research supervision.

References

All Tables

All Figures

	Fig. 1 Structure of flow stabilizing valve. 1. Valve cover 2. Ball valve spool 3. Valve sleeve 4. Seal ring 5. Annular groove 6. Valve core annular groove 7. Valve core 8. Retaining ring 9. Valve body 10. Oil-in 11. Core oil passage A 12. Oil source unloading port 13. Core oil passage B 14. Load unloading port 15. Liquid filling port 16. End cap 17. Load oil port.
In the text

	Fig. 2 System circuit with flow stabilizing valve.
In the text

	Fig. 3 Direct drive system simulation model of non-flow stabilizing valve. 1. Input signal 2. Motor 3. Pump 4. Safety valve 5. Safety valve 6. Hydraulic cylinder 7. Mass block 8. Disturbing force 9. Disturbance signal.
In the text

	Fig. 4 Direct drive electro hydraulic actuator simulation model with flow stabilizing valve. 1. Input signal 2. Motor 3. Pump 4. Check valve 5. Hydraulic cylinder 6. Check valve 7. Flow stabilizing valve 8. Flow stabilizing valve 9. Hydraulic cylinder 10. Mass block 11. Disturbing force 12. Disturbance signal.
In the text

	Fig. 5 Input signal of the system.
In the text

	Fig. 6 Disturbance signal of load.
In the text

	Fig. 7 Simulation curve of pulse disturbance signal with no flow stabilizing valve system.
In the text

	Fig. 8 Simulation curve of pulse disturbance signal with flow stabilizing valve system.
In the text

	Fig. 9 Velocity curve of step disturbance signal.
In the text

	Fig. 10 Velocity curve of slope disturbance signal.
In the text

	Fig. 11 Velocity curve of sine disturbance signal.
In the text

	Fig. 12 Chart of Hydraulic power part.
In the text

	Fig. 13 Chart of Electric control part.
In the text

	Fig. 14 Experimental response curve.
In the text