J. Cent. South Univ. (2020) 27: 3652-3666

DOI: https://doi.org/10.1007/s11771-020-4571-z

Pressure impact characteristic of vane type continuous rotary motor under different buffer structures

WANG Xiao-jing(王晓晶), HU Shan-liang(胡善良), SHEN Zhi-qi(沈志琦)

Key Laboratory of Advanced Manufacturing and Intelligent Technology of Ministry of Education,

Harbin University of Science and Technology, Harbin 150080, China

Central South University Press and Springer-Verlag GmbH Germany, part of Springer Nature 2020

Abstract: In order to solve the problem of pressure shock on the continuous rotary electro-hydraulic servo motor, the mathematical models of pressure gradient under the structure of pre-compressed chamber and U-shaped groove were established. The optimal structure dimensions of the pre-compressed chamber and the U-shaped groove were determined. The fluid models were established by Solidworks under the four structures of triangular groove, triangular groove with pre-compression chamber, U-shaped groove and U-shaped groove with pre-compression chamber. Simulation analysis of depressurization process of fluid models was performed based on FLUENT. The pressure nephograms of different buffer structures were compared and analyzed, and the pressure change curves and pressure gradient change curves in the process of depressurization were obtained. The results show that the optimal edge length of the pre-compressed chamber of continuous rotary electro-hydraulic servo motor is 20 mm in the process of decompression. The pressure reduction effect is the best when the width of the U-shaped groove is 1.5 mm and the depth is 1.65 mm. The U-shaped groove structure with pre-compression chamber is more conducive to alleviate the pressure shock phenomenon of the motor compared with different combine buffer structures.

Key words: continuous rotary electro-hydraulic servo motor; pressure impact; pre-compression chamber; U-shaped groove

Cite this article as: WANG Xiao-jing, HU Shan-liang, SHEN Zhi-qi. Pressure impact characteristic of vane type continuous rotary motor under different buffer structures [J]. Journal of Central South University, 2020, 27(12): 3652-3666. DOI: https://doi.org/10.1007/s11771-020-4571-z.

1 Introduction

The continuous rotary electro-hydraulic servo motor has many good performances, such as ultra-low speed, high frequency response, wide speed regulation and high precision, which is widely used to drive high-power simulation turntables, simulate various attitude angular movements, and reproduce various dynamic characteristics of its movements [1, 2]. The sudden change of pressure in the cavity is bound to huge pressure impact on the blade, due to the fact that the high and low-pressure chambers are switched instantaneously when the blades of the continuous rotary electro-hydraulic servo motor change direction. The top of the blade and the contour of the stator curve are worn when there is a large pressure impact and the blade expands frequently [3, 4]. On the other hand, pressure shock is accompanied by low-speed pulsation and flow pulsation, which seriously affects the low-speed performance of the continuous rotation electro- hydraulic servo motor.

The German IFD Institute conducted noise reduction research on the pressure shock phenomenon of hydraulic pumps and hydraulic motors [5, 6]. The elastic ring was placed in the elastic cavity around the cylinder ring. The energy was absorbed and released through the elastic ring when the elastic cavity was connected with the plunger cavity. Therefore, the pressure pulsation of the plunger pump was effectively reduced when the high and low pressure chambers were switched aiming at the pressure pulsation of plunger pump, PETTERSSON et al [7, 8] designed a valve plate with a pre-compression volume chamber, in which the plunger cavity was first connected to the pre-compression volume cavity, and the pressure of the plunger cavity was gradually compressed in the process of transition from oil suction to oil discharge. Thus, the pressure shock generated by the plunger pump in the process of rotation was reduced. The cycloid rotor pump was taken as the research object to study the effect of pressure relief groove by computational fluid dynamics method [9]. It is concluded that the pressure pulsation and the average pressure were reduced; meanwhile the torque difference between the inner and outer rotors was reduced when using the pressure relief groove. In addition, the potential damage of the cycloidal rotor pump was reduced due to the suppression of the cavitation and the reduction of the damage power of the cavitation [10]. The application of CFD analysis was introduced in the blade pump of non-invasive fluid wave aircraft engine control system. The goal was to understand the root cause of the pressure pulses in the system so as to propose improvements or corrective actions to the system. The effect of the change in leakage flow on the output pressure pulsation of the vane pump was studied to predict the output pressure fluctuation of the vane pump design stage [11]. The pressure pulsation was predicted theoretically by calculating the leakage flow and ideal flow. The calculated output pressure fluctuation waveform was in good agreement with the measured value. It was concluded that the change of leakage flow had a significant effect on the fluctuation of delivery pressure. OH et al [12] conducted quantitative analysis on the closed volume, outlet flow, pressure pulsation and cavitation phenomena of variable displacement vane pump through theoretical research. The analysis results show that the increase in temperature led to a slight decrease in exhaust flow and drive torque. The highest total efficiency was 73.8% at 80 °C and decreased by 3.5% at 120 °C. ZHANG et al [13] established a plunger pump calculation model based on AMESim software to analyze the dynamic characteristics of flow and pressure under different parameters, and concluded that the pulsation rate of flow and the pressure were inversely proportional to the tilt angle, rotation speed and outlet volume. LI et al [14] discussed the output pulsation and pressure pulsation of the vane pump under different working conditions through the fluid simulation software PUMPLINX and concluded that the output flow and pressure pulsation amplitude of the vane pump increased as the system pressure increased, LIU et al [15] performed modeling simulation and experimental tests on a new type of triangular buffer groove for the distribution plate structure, which provided a reference for the new plunger type hydraulic transformer to reduce pressure shock and noise. WU et al [16] compared the load pressure curve under the traditional grooved vane pump and the pre-compression grooved vane pump. They found that pre-compression grooved vane pump can effectively reduce the pulsation rate at the load inlet. Then they calculated the load pressure curves under different inclination angles of the distribution plate, and found that the minimum distribution plate inclination angle of outlet pressure fluctuation was 0.6° when the equivalent hole diameter was constant. The finite element simulation model of a plunger pump with a pre- compressed cavity was established by considering piston motion, fluid characteristics and leakage flow [17, 18]. Then the experiments were performed on the pump flow pulsation so as to obtain parameters of the pre- compressed cavity when the pressure shock was minimum.

There was less research on the pressure shock of continuous rotary electro-hydraulic servo motors, and the fluid simulation was rarely used to study the impact of different buffer structures on motor pressure shock. Therefore, in this paper, a theoretical analysis of the pressure reduction process of the sealed cavity of the continuous rotary electro-hydraulic servo motor was made. The mathematical models of pressure gradient in the pre-compression chamber and the U-shaped groove were established respectively, and the optimal structural sizes of the pre-compression chamber and U-shaped groove were determined. Then, the models were simulated and analyzed by FLUENT software, and the optimal structure was obtained by studying the impact of different buffer structures on the pressure shock of the motor. The article provided a new direction for doing the low speed performance of continuous rotary electro- hydraulic servo motor.

2 Establishment of mathematical model of pressure field in seal chamber

2.1 Principle of pre-compression chamber

During the working process of continuous rotary electro-hydraulic servo motor, a certain pressure shock is generated when the sealed container cavity connects with the waist groove of the valve plate. Therefore, the pre-compression chamber is introduced into the motor and connected to the sealed container cavity through a damping hole. The specific structure is shown in Figure 1. The oil inlet chamber of the motor is filled with high pressure oil, and the oil return chamber is filled with low pressure oil after starting the motor. With the rotation of the rotor, the inside of the sealed cavity is high-pressure oil when the blade is not in contact with the waist groove of the oil return cavity. The sealed chamber is connected to the oil return chamber, and the oil pressure in the sealed chamber gradually decreases when the blade is in contact with the waist groove of the oil return cavity. At this time, the pressure of the high- pressure oil in the pre-compression chamber is gradually released because of the damping hole. Therefore, the pressure shock phenomenon of the motor during the pressure reduction process is effectively improved, and the low-speed performance of the motor is also improved.

Figure 1 Pre-compressed chamber structure diagram

2.2 Pressure gradient mathematical model of pre-compression chamber

As shown in Figure 2, the research object is defined as the sealed cavity formed by the two adjacent blades of the motor, the flow plate, the rotor, the pre-compressed cavity and the internal curve of the stator. The change in oil pressure in the sealed cavity is obtained by using the following differential equation:

                            (1)

where dp is the differential of hydraulic pressure; β_e is the oil elasticity modulus (MPa); V is the initial volume of oil in sealed chamber (m³); dV is the volume differential of oil in sealed chamber.

Figure 2 Section of motor sealing cavity

As the continuous rotation of the motor, the oil in the seal cavity is gradually compressed. The pressure increment is positive and the volume increment is negative, so a negative sign is put in front of the equation.

Assuming that the initial volume of the sealed cavity is V, dθ/dt=ω, Eq. (1) can be transformed as follows:

                       (2)

The oil hydraulic pressure gradient in the sealing chamber is expressed as:

                            (3)

where θ is the rotor angle (°) and ω is the motor speed ((°)/s).

The volume change of the oil volume in the sealed cavity dV/dt is equivalent to the volume of liquid flowing into or out of the sealed cavity, including the volume change caused by the volume change and leakage of the buffer tank into (or out of) the sealed cavity. So, the pressure gradient is expressed as:

                      (4)

According to the flow rate formula of the orifice, the volume change from the buffer tank into the sealed cavity is obtained:

                      (5)

where C_q is the flow coefficient; △p is the pressure difference between the inlet and outlet; ρ is the density of the oil (kg/m³); A is the area of the orifice of the oil return cavity (m²).

Triangular buffer slots are provided at both ends of the four oil distribution windows. The structure of the triangular groove can be regarded as a triangular pyramid body machined by a molding tool. The plane angle of the buffer groove is φ=10°, and its shape and dimension are shown in Figure 3.

The throttling area A of triangular buffer groove is expressed as:

                    (6)

Figure 3 Diagram of triangular groove structure

where R is the leading edge radius of the oil distribution window (m); θ is the angle of rotation of the motor blade on the triangular groove (°).

From formulas (5) and (6), the volume change of the buffer tank into or out of the sealed cavity is expressed as:

          (7)

The physical structure diagram of the continuous rotary motor is shown in Figure 4. The continuous rotary electro-hydraulic servo motor is composed of the plate, stator, rotor, blade and end cover. The 13 blades are mounted in the rotor slots of the motor. The leakage amount can be approximately expressed as the leakage between the top of the blade and the inner surface of the stator, and the leakage between the end surface of the blade and the distribution plate in the studied sealed cavity. It is assumed that p_s is the oil source pressure, p₁ is the pressure of the high-pressure chamber, p₂ is the pressure of the low-pressure chamber, and p_g is the outlet pressure of the relief valve. Then, the gap of each part of the motor is regarded as a parallel plate gap, and the change of the volume of the seal cavity caused by the leakage is expressed as:

                (8)

where B is the axial width of stator, mm; ω is the rotational angular speed of the motor; △ is the height of clearance, mm; μ is the dynamic viscosity coefficient (Pa·s); l₁ is the gap length, mm; R₁ is the radius of the long radius arc of the inner curved surface of the stator, mm; l₂ is blade width, mm; r is the rotor radius, mm; and B₁ is the clearance width at the long radius arc of stator, mm.

Figure 4 Internal structure diagram of continuous rotary electro-hydraulic servo motor:

The volume of the seal chamber is expressed as:

   (9)

where R₁ is the radius of short radius arc of stator inner surface, mm; z is the number of blades; L is the side length of pre-compression chamber, mm; r₁ is the radius of damping hole, mm; and l is the length of damping hole, mm.

Therefore, the pressure gradient in the stator’s long radius sealed cavity is expressed as:

              (10)

According to the working principle of the pre-compression chamber, the pressure shock of the motor during the pressure increasing and decreasing is reduced, as long as the pre-compression chamber is located between the two waist-shaped grooves of the flow plate. Therefore, this paper only conducts numerical simulation and optimization of the structural size of the pre-compressed chamber, and does not analyze the impact of different positions of the pre-compressed volume on the pressure shock of the motor. According to the pressure gradient Eq. (10), in the study of the depressurization process of continuous rotary electro-hydraulic servo motor, the pressure change in the sealed cavity is mainly determined by the volume of the pre-compression chamber, and the volume of the pre-compression chamber is mainly determined by the side length when the structural size of the triangular buffer groove is determined. Therefore, the structure of the pre-compressed chamber is optimized in order to determine the value of the side length of the structure when the pressure reduction effect is the best. The relevant parameter values in the calculation process are shown in Table 1.

Substituting the relevant parameters, the curves of pressure change and pressure gradient of different side sizes are obtained by Matlab, as shown in Figure 5. Curves 1, 2, 3, 4 and 5 respectively correspond to the pressure change curves and pressure gradient curves of pressure relief when the side length of pre-compression chamber is 10, 15, 20, 25 and 30 mm.

Table 1 Parameters of numerical solution

It can be seen from Figure 5 that curves 4 and 5 depressurize slowly when the blade turns over 3°. At this time, the pressure of the sealing chamber has not yet reached the outlet pressure, so there is still a certain pressure impact when the sealing chamber contacts the oil return chamber. Curve 3 has the best pressure reduction effect, and the pressure in the sealed chamber is just the same as the outlet pressure. When the blade turns over 3°, the pressure decreases faster in curves 1 and 2.

The pressure in the seal chamber has already reached the outlet pressure when the blades turn over the buffer groove. But the pressure impact is large, which is not conducive to the low speed of the motor. In summary, the pressure reduction effect is the best when the side length of the pre- compressed chamber is 20 mm.

Figure 5 Pressure relief curves of sealing chamber:

2.3 Pressure gradient mathematical model of U-shaped groove

In order to form a closed working chamber between the two blades, the rotation angle of the blade on the U-shaped groove is the same as that of the triangular groove, both of which are 3°. Therefore, the specific shape of U-shaped groove is shown in Figure 6, and its structural dimensions are determined by the width b and the depth h of U-shaped groove.

Figure 6 U-shaped groove structure diagram:

As shown in Figure 6, the size of the flow passage area is determined by the longitudinal section between the U-shaped groove and the blade. In order to obtain the over current area of the U-shaped groove, the rotation angle of the blade relative to the semicircle tangent CD in the process of rotation is θ. The angle is α when the blade just turns to the semicircle,OG and EF are perpendicular to each other, so α can be expressed as:

                        (11)

where R is the leading edge radius of the oil distribution window (mm); b is the U-shaped groove width (mm).

Therefore, when the blade rotates to θ<α, the corresponding U-shaped groove over current area is expressed as:

         (12)

where h is the depth of the U-shaped groove (mm).

When the blade rotates to θ≥α, the corresponding U-shaped groove overcurrent area is expressed as:

A₂=bh                                 (13)

In summary, the U-shaped groove overcurrent area is expressed as:

  (14)

The pressure gradient equation in the stator long radius sealing cavity can be obtained by Eq. (15).

 (15)

During the pressure relief process of the continuous rotating electro-hydraulic servo motor, the rotation angle of the blade on the U-shaped groove is 3°. According to the specific size of the distribution plate and the position relationship of the U-shaped groove, the width b of the U-shaped groove is set equal to 1.5 mm at first, and the depth h is set as different values for simulation research. Then, the pressure change curve and pressure gradient change curve in the working chamber are obtained, as shown in Figure 7. The curves 1, 2, 3, 4 and 5 respectively correspond to the pressure change curves and pressure gradient curves when the depth of U-shaped groove is 1.35, 1.5, 1.65, 1.8 and 1.95 mm.

Figure 7 Pressure relief curves of sealing chamber with different depth of U-shaped groove:

It can be seen from Figure 7 that the pressure of curve 1 and 2 decreases slowly. The pressure of the seal chamber has not reached the outlet pressure when the blade rotates 3°, so there is still a certain pressure impact when the seal chamber contacts with the oil return chamber. Curve 3 has the best pressure reduction effect, and the pressure in the sealed chamber is just the same as the outlet pressure. The pressure decreasing of curves 4 and 5 is faster. The pressure in the seal chamber has already reached the outlet pressure when the blades rotate over the buffer groove. But the pressure impact is larger, which is not conducive to the low speed of the motor. Therefore, the pressure reduction effect is the best when the width of U-shaped groove is 1.5 mm and the depth is 1.65 mm. Then the depth h of the U-shaped groove is equal to 1.65 mm, and the width b is set to take different values for simulation research. The pressure change curves and pressure gradient change curves in the working chamber are obtained, as shown in Figure 8. The curves 1, 2, 3, 4 and 5 respectively correspond to the pressure change curves and pressure gradient curves when the widths of U-shaped groove are 1.4, 1.45, 1.5, 1.55 and 1.65 mm.

It can be seen from Figure 8 that the pressure reduction effect of curve 3 is the best. The pressure of the seal chamber is just the same as the outlet pressure when the blade rotates 3°, the transition is relatively stable, and the pressure gradient is moderate, which can meet the requirements of low-speed performance of the motor. Therefore, the U-shaped groove has the best effect of lowering pressure when the width and depth of the U-shaped groove are 1.5 and 1.65 mm, respectively, based on the analysis results in Figure 7.

Figure 8 Pressure relief curves of sealing chamber with different width of U-shaped groove:

In section 2, the working principle of pre- compression chamber is introduced, and the mathematical models of pressure gradient in pre- compression chamber and U-shaped groove are established.

3 Simulation and analysis of step-down process of continuous rotary electro- hydraulic servo motor

3.1 Pretreatment of fluid calculation

In order to study the influence of different buffer structures on the pressure impact during the motor’s decompressing process, the sealed chamber models of nine positions are established by using SolidWorks, in which two blades rotate from 0° to 4°at an interval of 0.5°. Figure 9 shows the fluid models under the four structures of the triangular groove with and without pre-compression chamber and the U-shaped groove with and without pre- compression chamber when the two blades rotates 2°. The dimensions of the pre-compressed chamber and the U-shaped groove are obtained through numerical optimization in the previous section.

Mesh generation is a key step in the pre- processing of flow field analysis, in which the quantity and quality of mesh determine the accuracy of the final calculation results. In general, the calculation accuracy will be improved to some extent with the increase of the number of grids. But the increase in the number of grids does not contribute significantly to the calculation accuracy for the regions with small model variation gradients, while the impact on the calculation cost is obvious. Therefore, the completed sealed cavity fluid models are meshed through the professional grid division software ICEM-CFD as shown in Figure 10 [19, 20]. The unstructured meshes are used because the gap between the top of the motor blade and the inner wall of the stator and the dimensions of triangular groove tip are small. In order to ensure the accuracy of calculation, create density function of ICEM-CFD is used to locally encrypt areas with sharp corners and significant pressure changes. Figure 10 shows the grid model after partitioning. It can be seen from Figure 11 that the grid quality is all greater than 0.3, which meets the calculation requirements.

Figure 9 Fluid models of sealed cavity:

Figure 10 Internal flow field mesh mode:

Figure 11 Grid quality decision diagram

The completed mesh models are imported into FLUENT [21, 22], by which the renormalization group (RNG) k-ε turbulence model is selected, the 32nd anti-wear hydraulic oil is used as the fluid medium; the oil density is 872 kg/m³, the oil viscosity is 0.0279 kg/(m·s), the inlet oil pressure is 7.5 MPa, and the outlet oil pressure is 2.5 MPa. Then the iterative steps are set for calculation.

The result is considered to be convergent when the general residual is reduced to 10^-3 in the process of numerical solution of FLUENT software, but the accuracy of the result cannot be determined directly according to the residual value. Because the residual value is the result obtained by taking the average value in the whole flow field calculation area, the variation trend of the residual with the number of iteration steps is not a strict standard to judge whether to converge. Moreover, parameter monitoring points should be set up inside the fluid model to monitor the changes of relevant parameters.

The calculation convergence of the motor during the pressure reduction process is judged by monitoring the residuals and sealing cavity pressure values. The pressure point monitoring curve of the sealed cavity derived from FLUENT is shown in Figure 12. The abscissa represents the number of iterations calculated numerically, and the ordinate is the residual value. The calculation results can be considered convergent when all parameters values of the residual curves are below 10^-3 and the pressure value of the monitoring point tends to be stable.

3.2 Fluid simulation of sealed cavity in triangular groove with and without pre- compression chamber

After the parameter setting is completed, all calculation results obtained by FLUENT are output to CFD-POST software for post-processing. The pressure nephograms of the triangular groove structure with and without the pre-compression chamber structure are obtained when the blade rotating angles of continuous rotary electro- hydraulic servo motor are 0°, 0.5°, 1°, 1.5°, 2°, 2.5°, 3°, 3.5° and 4°, as shown in Figures 13 and14.

Figure 12 Convergence monitoring curves:

In order to analyze the impact of different buffer structures on the pressure shock during the pressure reduction of the motor, the pressure value of the working chamber corresponding to each angle is extracted through the pressure nephograms. Then the change curves of the pressure in the working chamber with and without the pre-compression chamber respect to the rotation angle are shown in Figure 15(a). The pressure gradient curves of the working chamber of the motor are obtained by MATLAB, as shown in Figure 15(b).

In order to analyze the impact of different buffer structures on the pressure shock during the pressure reduction of the motor, the pressure value of the working chamber corresponding to each angle is extracted through the pressure nephograms. Then the change curves of the pressure in the working chamber with and without the pre-compression chamber with respect to the rotation angle are shown in Figure 15(a). The pressure gradient curves of the working chamber of the motor are obtained by MATLAB, as shown in Figure 15(b).

It can be seen from Figure 15(a) that there is a certain fluctuation in pressure during the pressure reduction process when the buffer structure is the triangular groove. But the pressure fluctuation decreased, and the whole depressurization process tended to be smooth after the introduction of the pre-compression chamber structure. It can be seen from Figure 15(b) that the pressure gradient in the sealed volume is reduced after the pre-compression chamber is added. Therefore, the pre-compression chamber respect can alleviate the pressure shock by the motor during the decompression process.

3.3 Fluid simulation of sealed cavity in U-shaped groove with and without pre-compression chamber

The calculation results obtained by FLUENT are output to the CFD-POST software for post- processing. The pressure nephograms of the U-shaped groove structure with and without the pre-compression chamber structure are obtained when the blade rotating angles of continuous rotary electro-hydraulic servo motor are 0°, 0.5°, 1°, 1.5°, 2°, 2.5°, 3°, 3.5° and 4°, as shown in Figures 16 and 17.

It can be seen from Figure 17 that the flow area of the U-shaped groove gradually increases with the continuous rotation of the motor blades, and the pressure of the seal chamber decreases to the pressure of the return chamber gradually. However, different from the previous buffer structures, the composite buffer structure combined with the U-shaped groove and the pre-compression chamber has no pressure fluctuation in the pressure cloud diagram, and the process of lowering pressure tends to be stable. The pressure value of the working chamber corresponding to each angle is extracted through the pressure nephograms. Then the change curves of the pressure in the working chamber respect to the rotation angle are shown in Figure 18(a). The pressure gradient curves of the working chamber of the motor are obtained by MATLAB, as shown in Figure 18(b).

Figure 13 Pre-compressed cavity structure diagram:

Figure 14 Pressure field distribution of fluid model with triangular groove and pre-compressed cavity:

Figure 15 Pressure relief curves of sealing chamber:

Figure 16 Pressure field distribution of fluid model under U-shaped groove structure:

It can be seen from Figure 18(a) that there is a certain fluctuation in pressure during the pressure reduction process when the buffer structure is the U-shaped groove. But after the introduction of the pre-compression chamber structure, the pressure fluctuation decreases, and the whole depressurization process tends to be smooth. It can be seen from Figure 18(b) that after the pre- compression chamber is added, the pressure gradient in the sealed volume is reduced. Therefore, the pre-compression chamber can alleviate the pressure shock produced by the motor during the decompression process. In addition, the composite buffer structure combined with the U-shaped groove and the pre-compression chamber has better lowering effect compared with the other three buffer structures.

Figure 17 Pressure field distribution of fluid model under U-shaped groove and pre-compressed cavity structure:

Figure 18 Pressure relief curves of sealing chamber:

In section 3, the step-down process of continuous rotating electro-hydraulic servo motor is simulated and analyzed.

4 Conclusions

1) The mathematical model of the pressure gradient in the pre-compressed chamber structure is established by analyzing the working principle of continuous rotary electro-hydraulic servo motor. Then the relevant parameters are substituted to analyze the pressure change curves and the pressure gradient change curves in the working cavity under different pre-compressed chamber sizes. Finally, it is determined that the pressure reduction effect is the best when the side length of the pre-compressed chamber is 20 mm.

2) The mathematical model of the pressure gradient in the U-shaped groove structure is established. Then, the pressure change curves and pressure gradient curves in the working chamber under different U-shaped groove sizes are analyzed. It is concluded that the U-shaped groove has the best effect when the width is 1.5 mm and the depth is 1.65 mm.

3) The pressure nephograms under the four structures of the triangular groove with and without a pre-compression chamber and the U-shaped groove with and without a pre-compression chamber are analyzed as well as the pressure change curves and pressure gradient curves of the decompression process. It is determined that the U-shaped groove with pre-compression chamber on the motor valve plate is the most favorable to relieve the pressure impact of the motor.

Contributors

The overarching research goals were developed by WANG Xiao-jing, HU Shan-liang and SHEN Zhi-qi. WANG Xiao-jing established the mathematical model of the sealed chamber under different buffer structures. HU Shan-liang and SHEN Zhi-qi established 3D model and conducted simulation analysis. WANG Xiao-jing, HU Shan- liang and SHEN Zhi-qi analyzed the calculated results. The initial draft of the manuscript was written by WANG Xiao-jing, HU Shan-liang and SHEN Zhi-qi. All authors replied to reviewers’ comments and revised the final version.

Conflict of interest

WANG Xiao-jing, HU Shan-liang and SHEN Zhi-qi declare that they have no conflict of interest.

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(Edited by FANG Jing-hua)

中文导读

叶片式连续回转马达不同缓冲结构的压力冲击特性研究

摘要：为了解决连续回转电液伺服马达压力冲击问题，分别建立预压缩容腔和U型槽结构下的压力梯度数学模型，确定了预压缩容腔及U型槽的最优结构尺寸；然后，应用Solidworks软件分别建立有无预压缩容腔的三角槽和有无预压缩容腔的U型槽这四种结构下的流体模型；基于FLUENT软件对流体模型的降压过程进行仿真分析，对比分析不同缓冲结构下的压力云图，得到降压过程的压力变化曲线和压力梯度变化曲线。结果表明，连续回转电液伺服马达在降压过程中预压缩容腔的最优边长为20 mm；U型槽宽度为1.5 mm，深度为1.65 mm时降压效果最好；相较于不同组合缓冲结构，有预压缩容腔的U型槽结构更有利于缓解马达的压力冲击现象。

关键词：连续回转电液伺服马达；压力冲击；预压缩容腔；U型槽

Foundation item: Project(51975164) supported by the National Natural Science Foundation of China; Project(201908230358) supported by the China Scholarship Council; Project(2019-KYYWF-0205) supported by the Fundamental Research Foundation for Universities of Heilongjiang Province, China

Received date: 2020-04-26; Accepted date: 2020-09-16

Corresponding author: WANG Xiao-jing, PhD, Professor; Tel: +86-18903669159; E-mail: hitwangxiaojing@163.com; ORCID: https://orcid.org/0000-0001-6285-9193