Jing Yao， Yuxin Yin， Zhaosheng Dong and Yuantong He

（1. School of Mechanical Engineering, Yanshan University, Qinhuangdao 066004, China；2. School of Mechanical Engineering, Nanjing Institute of Technology, Nanjing 211167, China）

Abstract: For an ultra-high-pressure hydraulic transmission system of a large-size hydraulic forging press (LHFP), a 70 MPa two-way proportional cartridge valve has been developed to improve the power weight ratio of the hydraulic forging press. In this study, a nominal diameter 25 mm(DN25) cartridge valve is taken as the research object. A longer concentric cylindrical annular gap is set to effectively prevent the ultra-high-pressure oil from flowing to the pilot stage and a seated valve structure is set to form the linear sealing zone in the closing state of the main valve port.Electric-displacement feedback is adopted to realize precise control of the main valve port flow and the features of this valve are investigated. In order to verify the strength and static and dynamic characteristics, the finite element model and a simulation model of the valve proposed above are built. There is a little deformation which does not affect the main valve spool movement, and the main valve port flow meets the design demands. Then, the prototype of DN25 70TPCV is manufactured and a ultra-high-pressure experimental platform is developed. The experimental results show that the DN25 70TPCV designed in this study has the advantage of fast response, high control precision, and low leakage, which can meet the requirements of LHFPs.

Key words: ultra-high-pressure；cartridge valve；large-size hydraulic forging press(LHFP)；proportional control；modelling

The demand for high-strength forged titanium/aluminum alloy has increased a lot during recent years, especially in fields such as aviation,aerospace, nuclear power, etc. This is due to the increasing requirements for the forces of large-size hydraulic forging presses (LHFPs)[1]. However,the increase of tonnage means larger actuator action areas, which leads to increasingly large sizes of LHFPs. Therefore, increasing the system pressure is a good solution to this problem. Ultrahigh-pressure systems with a working pressure up to 70 MPa have been successfully applied to 80 000-ton large hydraulic forging presses in China. The increase of pressure in hydraulic systems means that critical components, especially proportional control components, need to perform reliably under ultra-high-pressure and large-flow conditions.However, domestic research in this area is not mature enough to meet the application requirements, which greatly limits the development of industrial technology[2].

In addition to studies on proportional cartridge valves for forging hydraulic presses[3?4], proportional cartridge valves for mobile applications have also been studied in recent years[5]. Through mathematical modeling and simulations, the initial design parameters of cartridge valves have been verified, and the dynamic and static characteristics under different operating conditions have been optimized[6?8]. To analyze the stress and deformation distribution of the cartridge valve under different pressures, the finite element method has been used for analysis and calculation. The analysis results have laid a theoretical foundation for the structural optimization of two-way proportional cartridge valves[9?10]. By studying the design of proportional cartridge valves with different feedback forms, scholars get their corresponding characteristics in terms of valve properties[11?13].

The two-way cartridge valves involved in the above studies are used in applications where the pressure can be up to 31.5 MPa. As for ultrahigh-pressure equipment, ultra-high pressure waterjets have been applied in practice, which are pressurized by hydraulic cylinders[14?17]. As for ultra-high-pressure hydraulic valves, there have been studies on ultra-high-pressure relief valves,pneumatic pressure reducing valves and valve bodies[18?20]. These studies have improved the reliability of ultra-high-pressure hydraulic valves.However, due to the specificity of the application of ultra-high-pressure proportional cartridge valves, the research results are mainly concentrated on foreign manufacturers and there is a lack of study in the current published literature.Therefore, research and development are extremely difficult.

According to the special working conditions of LHFPs ’ hydraulic proportional control systems, DN25 70TPCV should meet the following requirements:

①Max working pressure: 70 MPa

②Flow at 0.5 MPa Pressure drop: ≥360 L/min(When the main valve spool is at max working position)

③Leakage flow at 70 MPa Pressure drop:≤1 L/min (When the main valve spool is at zero position)

④Flow hysteresis: ≤±3% (Relative to maximum working flow and expressed as a percentage)

⑤Flow nonlinearity: ≤±3% (Relative to maximum working flow and expressed as a percentage)

⑥Flow non-repeatability: ≤3% (Relative to maximum working flow and expressed as a percentage)

⑦Displacement step response time: ≤30 ms(The time that the main spool moves from the initial position to the command position for the first time)

⑧Displacement step peak overshoot: ≤3%

The current challenges in ultra-high-pressure systems include the sealing and the leakage,the opening and closing response, the strength of parts, and the deformation and the movement seizure of the main valve spool movement, which increase the difficulties of designing the sealing device, control mechanism, material selection,and valve structure.

The design of DN25 70TPCV is designed based on the above challenges. A longer concentric cylindrical annular gap is set to effectively prevent the ultra-high-pressure oil from flowing to the pilot stage. A seated valve structure is set to form the linear sealing zone in the closing state of the main valve port. A finite element model is built to verify that deformation is little enough that it does not affect the main valve spool movement. The design of proportional flow control and the research on the features of this valve are used to meet the requirements of control characteristics, which are verified by the simulation and experiment.

Thus, for the ultra-high-pressure hydraulic system of a LHFP, a 70 MPa two-way proportional cartridge valve has been developed to improve the power-weight ratio of the hydraulic forging press. Meanwhile, the design of proportional control is applied to improve displacement control accuracy and reduce the impact of pressure relief on LHFPs.

1 Design of DN25 70TPCV

The initial structure of DN25 70TPCV is designed according to the above requirements, as shown in Fig.1. It comprises the pilot stage and the main valve stage. The pilot stage comprises a pilot valve, pilot piston, and spring. The main valve stage comprises a main valve spool, main valve sleeve, and transition sleeve. There are other parts such as O-rings that also play an important role.

In order to provide reliable performance for the hydraulic proportional control system of LHFPs under ultra-high-pressure and large-flow conditions, the main features can be summarized as follows.

1.1 Pressure isolation

As shown in Fig.2, a concentric cylindrical annular gap, which is set between the inner hole of the transition sleeve and the piston rod, is formed after considering the requirements for strength, assembly and leakage.

The most special situations which may occur in the working process of DN25 TPCV are selected, and the following assumptions are proposed. First, the flow state of the oil in the gap is laminar flow. Second, the pilot piston rod is completely concentric with the inner hole of the transition sleeve. Next, the pressure of the upper cavity of the main valve spool is 70 MPa and the pressure of the lower cavity of the pilot piston is 0 MPa. Finally, the pilot piston rod moves upwards by 8 mm in 30 ms and takes 1.2 times the average speed as the maximum speed, which is 0.32 m/s.

Fig. 1 DN25 70TPCV structure diagram

Fig. 2 Pressure isolation

In order to measure the isolation effect, a calculation formula for the leakage of a concentric cylindrical annular gap is introduced as

whereQvis the leakage from the upper cavity of the main valve spool to the lower cavity of the pilot piston,dis the diameter of the piston rod,which is 7 mm,his the maximum thickness of the clearance between the inner hole of the transition sleeve and the piston rod, which is 0.01 mm,μis dynamic viscosity, which is 0.04 N·s/mm2(46# Hydraulic oil at 40 ℃),lis the length of the clearance, which is 23 mm, anduis the maximum upward movement speed of the piston rod,which is 0.32 m/s.

After calculation, the maximum leakage from the upper cavity of the main valve spool to the lower cavity of pilot piston is 0.1 L/min. This only accounts for 0.25% of the pilot valve’s rated flow, so it has less effect on the movement control of the pilot piston.

1.2 Seated valve structure

A seated valve structure is incorporated into the main valve stage for low leakage flow, as shown in Fig.3.

By designing 55 and 45 degree chamfers for the main spool and the sleeve at the main valve port area, respectively, the linear sealing zone is formed in the closing state of the main valve port. The spring is set on the upper cavity of the pilot stage to ensure that the main valve port of the main valve maintains reliably closed when the pilot valve is de-energized.

Fig. 3 Structure of the main valve port

1.3 New radial static seal

As shown in Fig.4, by using a one-piece radial seal ring in the fitting clearance between the main valve stage and the mounting hole to replace an O-ring and retaining ring, leakage under 70 MPa is prevented. To ensure the same sealing effect, the integral radial sealing ring is used for sealing, which increases tolerance and reduces manufacturing difficulty and assembly difficulty, but also increases cost.

Fig. 4 Radial static seal ring

1.4 Hollow main valve spool

As shown in Fig.5, the middle part of the main valve spool is provided with an axial blind hole and six small radial through holes for connecting port A with the upper cavity of the main valve spool.

As shown in Fig.5,FK,PAA1, mg, andfsalways hinder the valve from opening, andPAA2,PAA4, andPBA3always hinder the valve from closing.FBandfdalways impede the movement of the spool, andFpis determined by the control signal.

As shown in Fig.5a,Fphinders the valve from opening. In order to ensure that the main valve port is reliably closed, the following formula must be established

As shown in Fig.5b,Fphinders the valve from closing andFHhinders the valve from opening. The direction to help the valve close is defined as positive. In order to ensure the controllability of the main spool displacement during dynamics, and reduce the impact of load changes on the pilot control performance, the following formula must be established

whereλis the is the margin coefficient with a value of 1.2-1.4,mis the total mass of the moving parts,xis the displacement of main spool,A1is the active area of the upper cavity of the main valve spool,A2is the active area of the inner cavity’s bottom of the main spool when the main valve port is closed,A3is the annular active area outside the main valve spool when the main valve port is closed,A4the active area of the lower cavity of the main valve spool when the main valve port is open,PAis the pressure of port A,PBis the pressure of port B,Fpis the output force of the pilot stage, determined by the pilot’s position,FKis the output force of the spring at the top of the pilot piston,FBis the viscous damping force between the main valve spool and sleeve,FHis the hydraulic power at the main valve port,fsis the static friction between the main valve spool and sleeve, andfdis the dynamic friction between the main valve spool and sleeve.

1.5 Proportional flow control

The pilot piston of DN25 70TPCV is controlled by a Direct Drive Servo-Proportional Control Valve (pilot valve), and the LVDT installed at the top works as the feedback unit.Electric-displacement feedback is adopted to improve the displacement control accuracy of the pilot piston. The pilot rod is integrated into the main valve spool by screw thread, so the main valve spool is directly driven by the pilot piston to achieve precise control of the main valve port flow. The hydraulic principle diagram of DN25 70TPCV is shown in Fig.6. The displacement control block diagram of the main valve spool is shown in Fig.7.

Fig. 5 Stress diagram of main valve stage

Fig. 6 Hydraulic principle diagram

Fig. 7 Displacement control block diagram of the main valve spool

As shown in Fig.7, the displacement of the main spool is a closed loop controlled by the PID controller.Gf(s) is the transfer function of the external load applied to the system,G1(s) andG2(s) are the inherent transfer functions of DN25 70TPCV, andKXis the gain of the LVDT.

1.6 Flow-displacement characteristics

The bottom of the main valve spool is provided with a rectangular main valve port to ensure the linearity of the flow when the main valve port has a small opening. The main valve port is full circumference when it has a large opening. Thus, the linear and variable gain flowdisplacement characteristics of the main valve port are achieved. As a proportional flow control valve, the main valve port flow can be defined as whereQis the main valve port flow,Cdis the flow coefficient which is related to the actual Reynolds number,XVis the displacement of main valve spool, ΔPis the pressure drop across the main valve port of main valve stage,ρis the fluid density, andWis the wet perimeter, which is related toXVand the structure of the main valve port, whose relationship can be defined as

whereXdis the largest displacement when the main valve port is in the dead zone,Xris the largest displacement when the main valve port is rectangular,Xtis the largest displacement where the main valve port is full circumference.Xd,XrandXtare designed and optimized according to the requirements of main valve port flow characteristic, andXtis limited by mechanical structure.

2 Structural Analysis

In addition to achieving precise proportional control of the main valve port flow, it is also important to ensure that the main valve port is able to be reliably opened and closed at 70 MPa.However, the deformation of the main valve stage’s parts may cause larger leakage of the main valve port when it is closed and movement seizure of the main valve spool when it is moving. In order to solve this problem, we apply SOLIDWORKS Simulation because it has the advantages of simple operation, strong practicality, and easy modification of parameters in the model. Using its intuitive and fast simulation function, a finite element model (FEM) is established to analyze the deformation of the main valve stage’s parts to facilitate getting an early understanding of its performance and avoid costly and time-consuming design.

Since the rated working pressure of the main valve stage (70 MPa) is much larger than the pilot valve stage (14 MPa), the main valve spool,the main valve sleeve, and the transition sleeve of the main valve stage are selected as research objects. The structure of the main valve stage is approximately symmetrical with respect to the center axis of the main valve spool, so the axial section is selected as a two-dimensional model to simplify the analysis process. The mounting form of the main valve and the distribution of the pressure should be considered. Thus, a fixed constraint is set for the threaded part at the top of the transition sleeve and a roller constraint is set for the part where the transition sleeve and the main valve sleeve are fitted with the mounting hole. Then, a load (up to 70 MPa) is applied to the parts that are immersed in the hydraulic oil and the output force of the pilot rod is applied to the upper end of the main valve spool. The mesh is automatically divided by SOLIDWORKS Simulation based on the scaling factor. After multiple trials, the scaling factor is determined to be set to 4 and the mesh density is set to fine. With these parameters, the mesh calculation time is short, and the mesh calculation error is less than 5% if the mesh is further refined, so the simulation results are reliable.

Based on the above work, two typical working conditions are selected to analyze the stress and deformation of the main valve stage: the main valve port is closed or kept 50% open. The material and yield strength of the three parts during the simulation are shown in Tab.1.

Tab. 1 Material and main mechanical characteristic of three parts

As shown in Fig.8 under the condition that the main valve port is closed, when port A is under 70 MPa and port B is under 0 MPa of pressure, the largest deformation occurs at the bottom of the seal groove (20 μm). Here, the stress is 477.6 MPa and the maximum deformation of the clearance between the main valve spool and the main valve sleeve is 5 μm. When the main valve port is opened 50%, the largest deformation still occurs at the bottom of the seal groove(16.65 μm). Here, the stress is 465.3 MPa and the maximum deformation of the clearance between the main valve spool and the main valve sleeve is 9.83 μm. Since the static pressure state of the valve spool and the maximum deformation of the clearance between the main valve spool and the main valve sleeve remain unchanged when the main valve port is opened, the maximum deformation of the clearance between the main valve spool and the main valve sleeve remains 9.83 μm.

Fig. 8 FEM analysis nephogram

In the above two cases, the local maximum stress is 477.6 MPa, which is less than the yield strength of the material, so it will not cause damage to the parts. The maximum deformation of the clearance between the main valve spool and the main valve sleeve does not exceed 10 μm (the minimum unilateral fit clearance), so it will not cause the larger leakage of the main valve port and the movement seizure of the main valve spool movement. This ensures that the valve can work reliably under a pressure of 70 MPa.

3 Experiment Study

3.1 Prototype production

Through the deduction of the main features above and FEM analysis, the main dimensions and parameters of the valve are obtained, as shown in Tab.2. After refining the dimensions of machined parts and selecting the standard parts,the machining and assembly of DN25 70TPCV are completed. The prototype is shown in Fig.9.

Tab. 2 Dimensions and parameters of DN25 70TPCV

Fig. 9 Prototype of 70TSPCV

3.2 Experimental rig

The prototype is mounted on two experimental platforms respectively. Then the experiments related to dynamic, static, and ultra-highpressure characteristics are conducted in order.

The experimental platform for dynamic and static characteristics is shown in Fig.10a and the corresponding hydraulic schematic diagram of the main hydraulic components is shown in Fig.10b.The maximum pressure of the main circuit is 35 MPa and the maximum flow is 600 L/min.The experimental pressure of the pilot circuit is 15 MPa and the experimental flow is 40 L/min.It can adjust the pressure and flow continuously and maintain a constant differential pressure at the main valve port.

The ultra-high-pressure experimental platform is shown in Fig.11a, and the corresponding hydraulic schematic diagram is shown in Fig.11b.The maximum experimental pressure of the main circuit is 120 MPa and the maximum experimental flow is 6 L/min. It can adjust the pressure and flow continuously as well. Using this experimental platform, the pressure maintenance test for different time periods and the pressure impact test for different frequencies can be performed.

Fig. 10 Dynamic and static experimental platform

Fig. 11 Ultra-high-pressure experimental platform

The signal acquisition and control system for DN25 70TPCV is shown in Fig.12. Based on RMCTools, data Interaction between a PC and a Motion controller is achieved. The motion controller is responsible for signal acquisition and control of DN25 70TPCV. Through the I/O port,the pressure, flow and displacement signals from DN25 70TPCV can simultaneously be collected and processed by the real time controller under a 1 ms sample time. Moreover, the user interface can help monitor data and adjust system control parameters. In this way, in the displacement continuous control system, the displacement deviation can be used to drive the pilot valve to accomplish accurate closed-loop control.

3.3 Static and dynamic characteristics experiment results and analysis

3.3.4 Differential pressure-flow characteristic experiment

The step displacement signals of 2.465 mm,3.08 mm, 3.69 mm, 4.925 mm, and 8 mm (respectively corresponding to the main valve at 10%, 20%, 30%, 50%, and 100% open) are respectively given to the main valve spool. When the steady state is reached, change the flow at the main valve port and then measure the differential pressure. The differential pressure-flow curve is shown in Fig.16.

As seen in Fig.16, the smaller the main valve opening, the more insensitive the flow of the main valve port to the differential pressure.

Fig. 16 Differential pressure-flow curve

3.4 Ultra-high-pressure experiment results and analysis

This part of experiment is accomplished on the ultra-high-pressure experimental platform,and the working fluid is commercial ISO VG46 mineral oil.

The main valve port is kept closed by giving an open-loop voltage signal to pilot valve.Then, the pressure of port A is increased gradually from 0 with a gradient of 5 MPa. At each pressure level, the pressure is maintained for 5 min. The target pressure is 105 MPa (1.5 times the rated pressure). Due to the influence of the maximum experimental flow, the maximum pressure that can be reached during this experiment is 95 MPa. Limited by the measuring range of the flow sensor (0.1–6 L/min), when the leakage flow is less than 0.1 L/min, the volume of the leakage liquid is recorded with a measuring cup during the pressure maintaining time, and the leakage flow is measured by the flow sensor when the leakage is between 0.1 L/min and 6 L/min.The leakage flow test curve is shown in Fig.17.

Fig. 17 Leakage flow test curve

As seen in Fig.17, when the pressure at port A varies from 0 to 5 MPa, no leakage occurs at the main valve port. The leakage at 10 MPa is 1.5 ml/min and the leakage at 70 MPa is 0.3 L/min.This meets the design requirements. Above 70 MPa,the leakage changes rapidly with the increase of port A’s pressure; the leakage reaches 1.53 L/min at 95 MPa.

After the above experiment, in order to verify the anti-impact characteristic (The influence of flow and pressure impact on the main valve port in the closed state) of DN25 70TPCV, the pilot valve open-loop voltage signal is also given to keep the main valve port closed. Then the alternating pressures of 0 and 70 MPa are given to port A with the frequency of 1 Hz for 5 min. Experimental result show that no deformation occurs in the valve body during the process and there is no oil leakage at the interface between the pilot manifold and the mounting manifold.

Finally, the main valve spool is controlled to move sinusoidally (with a center of 4 mm, amplitude of 4 mm, and period of 2 s), and there is no stagnation of the main valve spool movement.This means DN25 70TPCV can work reliably under a pressure of 70 MPa.

4 Conclusion

Aiming to satisfy the challenges of structure strength and control characteristics of proportional cartridge valves used in ultra-high-pressure conditions, a novel structure of DN25 70TPCV is developed in this study to meet the application requirements of LHFP. The DN25 70TPCV simulation models and finite element model are established, and the DN25 prototype and the experimental platforms are made.The simulation and the experimental results prove that the average idle displacement step response time of DN25 70TPCV is 23.5 ms, the peak overshoot is 1.525%,the maximum flow of the main valve port is 428.07 L/min at a differential pressure of 0.5 MPa,the maximum flow hysteresis is 2.59%, and the maximum linearity is –1.67%. When the valve is closed, the leakage of the main valve port at 70 MPa is 0.3 L/min. All the above parameters meet the requirements of LHFP.