Prestage-based electro-hydraulic servo valve driving force increasing method and electro-hydraulic servo valve

Abstract

The invention belongs to the technical field of electro-hydraulic servo valves, and relates to a prestage-based electro-hydraulic servo valve driving force increasing method, which comprises the following steps of: configuring a benchmarking electro-hydraulic servo valve and a to-be-calibrated electro-hydraulic servo valve; respectively obtaining a benchmarking pressure-current characteristic curve of the benchmarking electro-hydraulic servo valve, a benchmarking pressure-displacement characteristic curve of the prestage, a pressure-current characteristic curve of the electro-hydraulic servo valve to be calibrated and a pressure-displacement characteristic curve of the prestage; under the constant value input current, the benchmarking displacement of the benchmarking electro-hydraulic servo valve deflection plate and the actual displacement of the electro-hydraulic servo valve deflection plate to be calibrated are obtained; the theoretical displacement of a deflection plate of the electro-hydraulic servo valve to be calibrated is calculated, the pressure-displacement characteristic curve is corrected, and the benchmarking displacement of the benchmarking electro-hydraulic servo valve under the constant input current is found out; and lifting the theoretical displacement to the benchmarking displacement based on a preset lifting scheme. The preset lifting scheme comprises the steps of improving the rigidity of the bourdon tube and the feedback rod, increasing the width of the jet orifice of the nozzle and reducing the width between the two receiving ports.

Description

TECHNICAL FIELD

[0001] The present application relates to the technical field of electro-hydraulic servo valve, and particularly relates to a method for improving driving force of an electro-hydraulic servo valve based on a pre-stage and an electro-hydraulic servo valve. BACKGROUND

[0002] The electro-hydraulic servo valve is a highly precise component combining mechanical, electronic and hydraulic technologies. Due to its excellent control accuracy, fast dynamic response, flexible electrical signal processing capability, powerful power amplification characteristics and compact structure design, it has become the "nerve center" of modern high-performance hydraulic systems. In key fields such as large transport aircraft and civil passenger aircraft, its performance directly determines the reliability and safety of various types of rudders, control systems and actuators, and is crucial to flight control.

[0003] However, the actual operating conditions of the electro-hydraulic servo valve are often extremely harsh and complex. The precise fit gap between the valve core and the valve sleeve (usually only a few microns) makes it extremely sensitive to pollutants in the working environment (such as solid particles, gum, water vapor, etc. in the oil). These pollutants are extremely easy to accumulate, jam or form sludge in the valve core-sleeve gap, generating significant stiction resistance. At the same time, severe temperature fluctuations (such as low temperature at high altitude, high temperature near the engine compartment or self-heating during high-power operation) will cause uneven thermal expansion / contraction deformation of different material components such as valve core, valve sleeve and support structure, further exacerbating the resistance of valve core movement and even causing complete jamming.

[0004] Currently, the main ideas of the industry to deal with the valve core jamming problem are as follows: first, improve the cleanliness of the oil, such as using more precise filtration systems and strictly controlling the oil pollution level; second, improve the material and surface treatment, such as selecting wear-resistant and corrosion-resistant materials and applying special surface coatings to reduce friction and adhesion of pollutants; third, optimize the structural design, such as improving the geometry of the valve core and sleeve and the design of the pressure balance groove to balance the radial force and reduce hydraulic clamping force; fourth, strengthen maintenance, such as specifying strict regular flushing, testing and replacement cycles. However, the above-mentioned solutions have some limitations. First, it is difficult to ensure absolute pollution control, because in actual complex operating conditions, risks always exist, and fine particles or soft pollutants may still invade the precise gap. Second, the inherent characteristics of temperature deformation are difficult to completely eliminate, and the difference in thermal expansion coefficient of different materials is an inherent physical phenomenon. Under extreme temperature gradients, the jamming force caused by deformation may be much larger than the design expectation. Third, the driving force of the traditional electro-hydraulic servo valve (usually provided by a torque motor or a force motor) is mainly designed to meet the dynamic response requirements under normal operating conditions, and the maximum output force / torque often has insufficient margin for the sudden and large jamming resistance generated under the above extreme adverse conditions. When the jamming resistance instantaneously exceeds the peak capacity of the driving mechanism, the valve core will be unable to move, resulting in system failure.

[0005] Therefore, there is an urgent need to propose a method that can significantly improve the output capability of the core drive unit of the electro-hydraulic servo valve, so that it can still provide sufficient driving torque or thrust when encountering abnormally large resistance caused by contaminants and temperature deformation, reliably overcome the jamming force caused by contamination, and ensure the normal movement and control function of the valve core. Summary of the Invention

[0006] The purpose of this invention is to address the shortcomings of existing technologies by proposing a method for increasing the driving force of an electro-hydraulic servo valve based on a pre-stage and an electro-hydraulic servo valve. By adjusting the stiffness of the spring tube, the stiffness of the feedback rod, the width of the nozzle jet orifice, and the width between the two receiving ports, the driving force can be increased from 35N to 54N, effectively overcoming the clamping force caused by contamination.

[0007] To achieve the above objectives, the present invention adopts the following technical solution: In a first aspect, the present invention provides a method for increasing the driving force of an electro-hydraulic servo valve based on a pre-stage, comprising the following steps: Configure a benchmark electro-hydraulic servo valve and a calibrated electro-hydraulic servo valve whose driving force needs to be increased; Obtain the benchmark pressure-current characteristic curve and the benchmark pressure-displacement characteristic curve of the front stage corresponding to the benchmark electro-hydraulic servo valve, and obtain the pressure-current characteristic curve and the pressure-displacement characteristic curve of the front stage corresponding to the electro-hydraulic servo valve to be calibrated. Based on the benchmark pressure-current characteristic curve and the benchmark pressure-displacement characteristic curve of the pre-stage, the benchmark displacement of the deflection plate of the benchmark electro-hydraulic servo valve under constant input current is obtained; based on the pressure-current characteristic curve and the pressure-displacement characteristic curve of the pre-stage, the actual displacement of the deflection plate of the electro-hydraulic servo valve to be calibrated under constant input current is obtained. The theoretical displacement of the deflection plate of the electro-hydraulic servo valve to be calibrated under constant input current is calculated, and the pressure-displacement characteristic curve is further corrected by the deviation between the theoretical displacement and the actual displacement until the deviation between the actual displacement and the theoretical displacement is within the preset deviation range, so as to obtain the corrected pressure-displacement characteristic curve. Find the benchmark displacement of the benchmark electro-hydraulic servo valve under constant input current on the corrected pressure-displacement characteristic curve; The constraints are as follows: the open-loop transfer function gain of the electro-hydraulic servo valve from input current to output flow remains constant; the bandwidth variation of the electro-hydraulic servo valve is within a preset range. Based on the constraints, the theoretical displacement of the deflection plate of the electro-hydraulic servo valve to be calibrated under constant input current is increased to the benchmark displacement of the electro-hydraulic servo valve under constant input current according to the preset driving force enhancement scheme; the preset driving force enhancement scheme includes: increasing the stiffness of the spring tube and the stiffness of the feedback rod, increasing the width of the nozzle jet orifice, and decreasing the width between the two receiving ports.

[0008] As one possible implementation, the driving force is increased from 35N to 54N, the stiffness of the Bourdon tube is increased to 12.9Nm / rad, the stiffness of the feedback rod is increased to 1510 N / m, the nozzle jet width is increased to 0.163mm, and the width between the two receiving ports is reduced to 0.56mm.

[0009] As one possible implementation approach, a pre-defined motivation enhancement scheme can be developed in the following manner: Construct a positive boosting model for the driving force of the electro-hydraulic servo valve to be calibrated: in, As the driving force; The effective area at both ends of the valve core; For preamplifier pressure gain; The distance from the rotation center of the armature baffle assembly to the center of the deflection plate; This refers to the electromagnetic torque coefficient of a permanent magnet torque motor. For the control current of the input control coil; The load torque generated by the fluid force of the pre-stage jet on the deflector plate is approximately zero; The moment of inertia of the armature assembly; For the complex field variable; is the viscous damping coefficient of the armature assembly; For the stiffness of the Bourdon tube; For the stiffness of the magnetic spring in a permanent magnet torque motor; For the feedback rod stiffness; The distance from the center of the deflection plate to the center of the feedback lever ball; The lifting parameters were determined based on the principle of not altering the overall structure of the armature assembly and the magnetic reluctance and magnetic flux of the torque motor stage. These lifting parameters included the stiffness of the Bourdon tube. Feedback rod stiffness Nozzle jet width and the width between the two receiving ports ; Configure constraints: in, Indicates the cross-sectional area of ​​the valve core; The net stiffness of the torque motor for the target electro-hydraulic servo valve; For the feedback rod stiffness; The distance from the center of the deflection plate to the center of the feedback lever ball; The distance from the rotation center of the armature baffle assembly to the center of the deflection plate; For pre-stage flow gain; The amplitude and bandwidth of the target electro-hydraulic servo valve; Minimum amplitude bandwidth for the target electro-hydraulic servo valve; For the target electro-hydraulic servo valve, the maximum amplitude bandwidth is required. is the open-loop amplification factor of the force feedback loop.

[0010] As one possible implementation, the pre-stage pressure gain is related to the nozzle jet width. The following relationship must be satisfied: Preamplifier pressure gain and the width between the two receivers The following relationship must be satisfied: Pre-stage flow gain and nozzle jet width The following relationship must be satisfied: Preamplifier flow gain and the width between the two receiver ports The following relationship must be satisfied: In the formula, This indicates the pressure gain of the preamplifier stage. This indicates the pre-stage flow gain. Indicates the nozzle jet width. This indicates the width between the two receiving ports.

[0011] As one possible implementation method, the pressure-displacement characteristic curves of the benchmark electro-hydraulic servo valve and the pre-stage of the electro-hydraulic servo valve to be calibrated with the increased driving force are obtained by the following method: By setting different valve core displacements, the pressure difference between the two receiving ports corresponding to each displacement is obtained, and the pressure-displacement characteristic curve of the pre-stage is obtained by plotting points.

[0012] As one possible implementation, the pressure-current characteristic curves of the benchmark electro-hydraulic servo valve and the electro-hydraulic servo valve to be calibrated with the increased driving force are obtained by the following method: By setting different input currents, the pressure difference between the two receiving ports corresponding to each input current is obtained, and the pressure-current characteristic curve is obtained by plotting points.

[0013] As one possible implementation, the theoretical displacement is denoted as... It is obtained by calculation as follows: The distance from the rotation center of the armature baffle assembly to the center of the deflection plate; For the armature corner.

[0014] As one possible implementation method, armature corner It is obtained through the following calculation: For the overall stiffness of the torque motor; This refers to the mechanical damping ratio of the torque motor. For the complex field variable; The natural frequency of the armature baffle assembly; This refers to the electromagnetic torque coefficient of a permanent magnet torque motor. For the control current of the input control coil; The distance from the rotation center of the armature baffle assembly to the center of the deflection plate; The distance from the center of the deflection plate to the center of the feedback lever ball; For the feedback rod stiffness; For valve core displacement; The load torque generated by the fluid force of the pre-stage jet on the deflector plate is approximately zero.

[0015] Secondly, the present invention provides an electro-hydraulic servo valve, which uses the pre-stage-based electro-hydraulic servo valve driving force enhancement method provided in the first aspect to increase the driving force from 35N to 54N in order to overcome the clamping force caused by contamination.

[0016] Compared with the prior art, the beneficial effects of the present invention are as follows: 1. The proposed method for increasing the driving force of an electro-hydraulic servo valve based on a pre-stage can increase the driving force from 35N to 54N by adjusting the stiffness of the spring tube, the stiffness of the feedback rod, the width of the nozzle jet orifice, and the width between the two receiving ports, effectively overcoming the clamping force caused by contamination.

[0017] 2. The proposed method for enhancing the driving force of an electro-hydraulic servo valve based on a pre-stage has formulated an optimal solution for enhancing the driving force of the servo valve. It can increase the driving force of the valve core while ensuring that the flow gain of the servo valve remains basically unchanged, and the adjusted parameters are all within a reasonable range.

[0018] 3. The proposed method for enhancing the driving force of an electro-hydraulic servo valve based on a pre-stage is to construct a positive enhancement model for the driving force of the electro-hydraulic servo valve to be calibrated. The enhancement parameters are determined based on the principle of not changing the overall structure of the armature assembly and the magnetic reluctance and magnetic flux of the torque motor stage, so that the driving force of the electro-hydraulic servo valve to be calibrated is enhanced to a level comparable to that of the benchmark electro-hydraulic servo valve. The solution has high feasibility. Attached Figure Description

[0019] The accompanying drawings, which are included to provide a further understanding of the invention and form part of this invention, illustrate exemplary embodiments of the invention and are used to explain the invention, but do not constitute an undue limitation of the invention. In the drawings: Figure 1 This is a schematic diagram of the electro-hydraulic servo valve structure in an embodiment of the present invention; Figure 2 This is a schematic diagram of the pre-stage structure of the electro-hydraulic servo valve in an embodiment of the present invention; Figure 3 This is a schematic diagram of the structural dimensions of the pre-stage of the electro-hydraulic servo valve in an embodiment of the present invention; Figure 4 This is a flowchart of the electro-hydraulic servo valve driving force enhancement method based on the pre-stage in an embodiment of the present invention; Figure 5 This is a schematic diagram of the turbulent jet flow field structure at the oil outlet of the electro-hydraulic servo valve jet disk in an embodiment of the present invention; Figure 6 This is a schematic diagram illustrating the relationship between the nozzle jet width, the receiver width, and the driving force in an embodiment of the present invention. Figure 7 These are the pressure characteristic curves and flow characteristic curves obtained through Ansys simulation in this embodiment of the invention.

[0020] Figure Labels 1-Coil, 2-Armature assembly, 3-Right receiving hole, 4-Bourdon tube, 5-V-groove, 6-Feedback rod, 7-Right load chamber, 8-Valve core, 9-Right control chamber, 10-Upper magnetic conductor, 11-Lower magnetic conductor, 12-Left receiving hole, 13-Deflection plate, 14-Jet disk, 15-Filter screen, 16-Left load chamber, 17-Left control chamber, 18-Valve sleeve, 19-Base, 20-V-groove. Detailed Implementation

[0021] To facilitate a clear description of the technical solutions in the embodiments of the present invention, the terms "first" and "second" are used to distinguish identical or similar items with essentially the same function and effect. For example, the first threshold and the second threshold are merely used to distinguish different thresholds and do not limit their order. Those skilled in the art will understand that the terms "first" and "second" do not limit the quantity or execution order, and that the terms "first" and "second" are not necessarily different.

[0022] It should be noted that in this invention, the terms "exemplary" or "for example" are used to indicate examples, illustrations, or descriptions. Any embodiment or design described as "exemplary" or "for example" in this invention should not be construed as being more preferred or advantageous than other embodiments or designs. Specifically, the use of terms such as "exemplary" or "for example" is intended to present the relevant concepts in a concrete manner.

[0023] In this invention, "at least one" refers to one or more, and "more than one" refers to two or more. "And / or" describes the relationship between related objects, indicating that three relationships can exist. For example, A and / or B can represent: A alone, A and B simultaneously, or B alone, where A and B can be singular or plural. The character " / " generally indicates that the preceding and following related objects are in an "or" relationship. "At least one" or similar expressions refer to any combination of these items, including any combination of singular or plural items. For example, "at least one of a, b, or c" can represent: a, b, c, a combination of a and b, a combination of a and c, a combination of b and c, or a, b, and c, where a, b, and c can be single or multiple.

[0024] The present invention aims to provide a method for increasing the driving force of an electro-hydraulic servo valve based on a pre-stage and an electro-hydraulic servo valve. By adjusting the stiffness of the spring tube, the stiffness of the feedback rod, the width of the nozzle jet orifice, and the width between the two receiving ports, the driving force can be increased from 35N to 54N, effectively overcoming the clamping force caused by contamination.

[0025] In a first aspect, embodiments of the present invention provide a method for increasing the driving force of an electro-hydraulic servo valve based on a pre-stage. By employing a pre-stage lifting scheme, the stiffness of the Bourdon tube and the feedback rod is increased, the nozzle jet orifice width is enlarged, and the width between the two receiving ports is reduced. This increases the theoretical displacement of the deflection plate of the electro-hydraulic servo valve to be calibrated under a constant input current to the benchmark displacement of the electro-hydraulic servo valve under a constant input current, thereby overcoming the clamping force caused by contamination and ensuring the normal movement and control function of the valve core.

[0026] See Figure 1The electro-hydraulic servo valve includes: coil 1, armature assembly 2, right receiving port 3, Bourdon tube 4, V-groove 5, feedback rod 6, right load chamber 7, valve core 8, right control chamber 9, upper magnetic conductor 10, lower magnetic conductor 11, left receiving port 12, deflection plate 13, jet plate 14, filter screen 15, left load chamber 16, left control chamber 17, valve sleeve 18, and base 19. The armature assembly 2 and Bourdon tube 4 form a T-shaped frame. When the torque motor of the electro-hydraulic servo valve operates, it first magnetizes the lower magnetic conductor 11 to polarize it. The DC current through coil 1 causes a change in the magnetic force in the diagonal air gap, and the magnitude of the current is proportional to the rotation angle of the armature assembly 2. The armature assembly 2, Bourdon tube 4, and feedback rod 6 are rigidly connected and supported by the thin wall of Bourdon tube 4. Oil enters the jet plate through the oil supply port and the guide port of the feedback rod 6. The oscillation of the armature assembly 2 causes a change in the position of the guide port on the feedback rod 6 relative to the receiving port of the jet disk 14, thereby generating a pressure difference between the left control chamber 17 and the right control chamber 9 at the outlet of the jet amplifier receiver of the deflection plate 13. When the feedback mechanism operates, the current in the coil 1 generates a magnetic force at the end of the armature assembly 2. The armature assembly 2 rotates under the action of the magnetic force, and drives the feedback rod 6 to rotate under the support of the spring tube 4. The position of the guide port relative to the receiving port of the jet disk 14 changes, causing the jet to flow to one end of the valve core 8, generating a pressure difference between the two ends of the valve core 8. The valve core 8 moves under the action of the pressure difference, and at the same time, the oil supply Ps is connected to the control chamber, and the other chamber is connected to the oil return Pr. The valve core 8 pushes the small ball at the end of the feedback rod 6, establishing a feedback torque on the armature assembly 2. When the feedback torque is balanced with the torque generated by the current, the valve core 8 stops at a certain position. The displacement of the valve core 8 is proportional to the input current. Under a certain pressure, the flow rate to the load is proportional to the displacement of the valve core 8.

[0027] See Figure 2 The pre-stage of the electro-hydraulic servo valve is a deflector jet type, consisting of a jet plate 14 and a deflector plate 13. The deflector plate 13 is located on the feedback rod 6, and the small ball at the end of the feedback rod 6 is inserted into the middle of the spool 8 of the slide valve stage. When there is no control current input to the torque motor, the deflector plate 13 is in the neutral position, the recovery pressure of the left and right receiving holes is equal, and the spool 8 has no displacement. When there is current input to the torque motor, the armature assembly 2 drives the feedback rod 6 to deflect, and the position of the deflector plate 13 on the feedback rod 6 changes, thereby causing different recovery pressures in the left and right receiving holes. This creates a pressure difference across the spool 8, causing the spool 8 to displace, and correspondingly generating a certain flow rate.

[0028] This embodiment uses, as follows Figure 3The following description uses a fuel-deflecting jet servo valve with the shown front-stage structure dimensions as an example. The main dimensions of the front-stage structure are as follows: the width of the primary jet port on the jet plate is 0.155mm; the width of the secondary jet port on the V-groove is 0.16mm; the angle of the V-groove is 36.7°; the distance from the upper end face of the V-groove to the primary jet port of the jet plate is 0.2mm; the distance from the lower end face of the V-groove to the receiving port of the jet plate is 0.2mm; and the width of the wedge at the receiving port on the jet plate is 0.1mm.

[0029] See Figure 4 The driving force enhancement method provided in this embodiment includes the following steps: Configure a benchmark electro-hydraulic servo valve and a calibrated electro-hydraulic servo valve whose driving force needs to be increased; The driving force enhancement method provided in this embodiment is to enhance the driving force of the electro-hydraulic servo valve to be calibrated to a level comparable to that of the benchmark electro-hydraulic servo valve.

[0030] Obtain the benchmark pressure-current characteristic curve and the benchmark pressure-displacement characteristic curve of the front stage corresponding to the benchmark electro-hydraulic servo valve, and obtain the pressure-current characteristic curve and the pressure-displacement characteristic curve of the front stage corresponding to the electro-hydraulic servo valve to be calibrated. As one possible implementation method, the pressure-displacement characteristic curves of the benchmark electro-hydraulic servo valve and the pre-stage of the electro-hydraulic servo valve to be calibrated with the increased driving force are obtained by the following method: By setting different valve core displacements, the pressure difference between the two receiving ports corresponding to each displacement is obtained, and the pressure-displacement characteristic curve of the pre-stage is obtained by plotting points.

[0031] As an example, by using flow field simulation software to set different valve core displacements, the pressure difference between the two receiving ports corresponding to each displacement is obtained, and the pressure-displacement characteristic curve of the pre-stage is obtained by plotting points.

[0032] As another possible implementation, the pressure-current characteristic curves of the benchmark electro-hydraulic servo valve and the electro-hydraulic servo valve to be calibrated with the increased driving force are obtained by the following method: By setting different input currents, the pressure difference between the two receiving ports corresponding to each input current is obtained, and the pressure-current characteristic curve is obtained by plotting points. Based on the benchmark pressure-current characteristic curve and the benchmark pressure-displacement characteristic curve of the pre-stage, the benchmark displacement of the deflection plate of the benchmark electro-hydraulic servo valve under constant input current is obtained; based on the pressure-current characteristic curve and the pressure-displacement characteristic curve of the pre-stage, the actual displacement of the deflection plate of the electro-hydraulic servo valve to be calibrated under constant input current is obtained. The theoretical displacement of the deflection plate of the electro-hydraulic servo valve to be calibrated under constant input current is calculated, and the pressure-displacement characteristic curve is further corrected by the deviation between the theoretical displacement and the actual displacement until the deviation between the actual displacement and the theoretical displacement is within the preset deviation range, so as to obtain the corrected pressure-displacement characteristic curve. As an example, based on the torque balance equation, the constant input current of the benchmark electro-hydraulic servo valve is calculated. calibrated displacement of the lower deflection plate And the electro-hydraulic servo valve to be calibrated at a constant input current. Actual displacement of the lower deflection plate The constant input current of the electro-hydraulic servo valve to be calibrated is calculated. Theoretical displacement of the lower deflection plate Comparison with theoretical displacement The actual displacement of the electro-hydraulic servo valve to be calibrated Based on the deviation between the two, the pressure-displacement characteristic curve is corrected, and the benchmark electro-hydraulic servo valve is found on the curve at a constant input current. The corrective value below .

[0033] As one possible implementation, the theoretical displacement is denoted as... It is obtained by calculation as follows: (1) The distance from the rotation center of the armature baffle assembly to the center of the deflection plate; For the armature corner.

[0034] As one possible implementation method, armature corner It is obtained through the following calculation: (2) For the overall stiffness of the torque motor; This refers to the mechanical damping ratio of the torque motor. For the complex field variable; The natural frequency of the armature baffle assembly; This refers to the electromagnetic torque coefficient of a permanent magnet torque motor. For the control current of the input control coil; The distance from the rotation center of the armature baffle assembly to the center of the deflection plate; The distance from the center of the deflection plate to the center of the feedback lever ball; For the feedback rod stiffness; For valve core displacement; The load torque generated by the fluid force of the pre-stage jet on the deflector plate is approximately zero.

[0035] Find the benchmark displacement of the benchmark electro-hydraulic servo valve under constant input current on the corrected pressure-displacement characteristic curve; The constraints are as follows: the open-loop transfer function gain of the electro-hydraulic servo valve from input current to output flow remains constant; the bandwidth variation of the electro-hydraulic servo valve is within a preset range. Based on the constraints, the theoretical displacement of the deflection plate of the electro-hydraulic servo valve to be calibrated under constant input current is increased to the benchmark displacement of the electro-hydraulic servo valve under constant input current according to the preset driving force enhancement scheme. The preset driving force enhancement scheme includes: increasing the stiffness of the spring tube and the stiffness of the feedback rod, increasing the width of the nozzle jet orifice, and decreasing the width between the two receiving ports.

[0036] As one possible implementation approach, a pre-defined motivation enhancement scheme can be developed in the following manner: Construct a positive boosting model for the driving force of the electro-hydraulic servo valve to be calibrated: (3) in, As the driving force; The effective area at both ends of the valve core; For preamplifier pressure gain; The distance from the rotation center of the armature baffle assembly to the center of the deflection plate; This refers to the electromagnetic torque coefficient of a permanent magnet torque motor. For the control current of the input control coil; The load torque generated by the fluid force of the pre-stage jet on the deflector plate is approximately zero; The moment of inertia of the armature assembly; For the complex field variable; is the viscous damping coefficient of the armature assembly; For the stiffness of the Bourdon tube; For the stiffness of the magnetic spring in a permanent magnet torque motor; For the feedback rod stiffness; The distance from the center of the deflection plate to the center of the feedback lever ball; The lifting parameters were determined based on the principle of not altering the overall structure of the armature assembly and the magnetic reluctance and magnetic flux of the torque motor stage. These lifting parameters included the stiffness of the Bourdon tube. Feedback rod stiffness Nozzle jet width and the width between the two receiving ports ; Configure constraints: (4) (5) in, Indicates the cross-sectional area of ​​the valve core; The net stiffness of the torque motor for the target electro-hydraulic servo valve; For the feedback rod stiffness; The distance from the center of the deflection plate to the center of the feedback lever ball; The distance from the rotation center of the armature baffle assembly to the center of the deflection plate; For pre-stage flow gain; The amplitude and bandwidth of the target electro-hydraulic servo valve; Minimum amplitude bandwidth for the target electro-hydraulic servo valve; For the target electro-hydraulic servo valve, the maximum amplitude bandwidth is required. is the open-loop amplification factor of the force feedback loop.

[0037] Turbulent jet structure at the oil outlet of the jet disk, such as Figure 5 As shown, the flow field structure is divided into two parts: an initial section and a main section. The initial section contains two regions: the core region's fluid velocity is unaffected by the surrounding static fluid and maintains its original outlet velocity; the velocity distribution in the mixing layer is similar to that in the main section. The distance between the virtual source and the jet inlet is ignored.

[0038] The initial unit width momentum exiting the orifice can be expressed as: (6) In the formula, This represents the initial momentum of the jet. Indicates the initial velocity of the jet. The density of the hydraulic oil, Let be the initial distance from the center of the deflector plate to the center of the feedback rod ball. Since the deflector plate is placed horizontally, the potential energy of the fluid before and after the jet remains unchanged. Under steady flow and the influence of gravity, the pressure energy of the incompressible viscous fluid is converted into kinetic energy. The ideal Bernoulli equation for the fluid along the streamline is: (7) In the formula, This represents the pressure change before and after the jet, which is a constant value of 21 MPa. The density of the hydraulic oil, This indicates the initial velocity of the jet.

[0039] Therefore, As a constant, the initial unit width momentum of the nozzle in equation (6) changes only with the nozzle orifice width. When the nozzle orifice width increases... Increases, thus increasing the initial unit width momentum. As the nozzle width increases, the total kinetic energy at the jet orifice also increases, resulting in greater kinetic energy reaching the receiver. Since all the kinetic energy of the jet is converted into pressure energy at the receiver, increasing the nozzle orifice width, while keeping other parameters constant, increases the pressure difference at the pre-stage receiver, thereby increasing the driving force.

[0040] Because the servo valve has a symmetrical structure, the pressure at the receiving port is always the same when the deflector plate is in the neutral position: (8) In the formula, This refers to the pressure at the receiving port when the deflection plate is in the neutral position. The width of the receiving port, The density of the hydraulic oil, This refers to the kinetic energy of the fluid at the receiving port.

[0041] When the width of the two receiving ports decreases, it is equivalent to reducing the width of the two receiving ports. From equation (8), it can be seen that when As the width of the two receiving ports decreases, the pressure at the receiving ports increases. Therefore, keeping other parameters constant, a smaller width at the two receiving ports leads to a greater pressure difference across the ports, which in turn increases the driving force.

[0042] As one possible implementation, simulations yielded the pre-stage pressure gain and the nozzle jet width. The following relationship must be satisfied: (9) Preamplifier pressure gain and the width between the two receivers The following relationship must be satisfied: (10) Pre-stage flow gain and nozzle jet width The following relationship must be satisfied: (11) Preamplifier flow gain and the width between the two receiver ports The following relationship must be satisfied: (12) In the formula, This indicates the pressure gain of the preamplifier stage. This indicates the pre-stage flow gain. Indicates the nozzle jet width. This indicates the width between the two receiving ports.

[0043] Based on the positive boosting model of the driving force of the electro-hydraulic servo valve to be calibrated, the constraints, the relationship between the pre-stage pressure gain and the nozzle jet width and the width between the two receiving ports, and the relationship between the pre-stage flow gain and the nozzle jet width and the width between the two receiving ports, the relationship between the pre-stage structural parameters and the driving force is obtained as follows: Figure 6 As shown.

[0044] While ensuring an increase in the overall valve's driving force, the valve's flow characteristics must not be altered. Therefore, based on the constraint expression, the feedback rod stiffness, Bourdon tube stiffness, and pre-stage flow gain coefficient are proportional. To maintain a relatively constant flow gain for the servo valve, adjusting the pre-stage nozzle orifice width and receiver width to increase the driving force will also increase the pre-stage flow gain; therefore, the stiffness of the feedback rod and Bourdon tube should be appropriately increased. For example, the maximum value of the driving force within the required parameter range is calculated using MATLAB, and the specific parameters are shown in Table 1. Table 1 Optimization Parameter Table A fluid model was established with a nozzle orifice width of 0.163 mm and a receiver orifice width of 0.56 mm. Pressure and flow characteristic curves were obtained through Ansys simulation, as shown below. Figure 7 As shown, the pressure gain and flow gain are 1.91 × 10⁻⁶. 10 Substituting 0.03724 into the driving force formula, we get a driving force of 54N.

[0045] As one possible implementation, the driving force is increased from 35N to 54N, the stiffness of the Bourdon tube is increased to 12.9Nm / rad, the stiffness of the feedback rod is increased to 1510 N / m, the nozzle jet width is increased to 0.163mm, and the width between the two receiving ports is reduced to 0.56mm.

[0046] For example, see Table 2, which shows the pre-stage structure parameters before and after optimization.

[0047] Table 2 Pre-stage structural parameters before and after optimization As shown in Table 2, by optimizing the structural parameters of the pre-stage, the driving force can be increased from 35N to 54N.

[0048] Secondly, embodiments of the present invention provide an electro-hydraulic servo valve, which uses the pre-stage-based electro-hydraulic servo valve driving force enhancement method provided in the first aspect to increase the driving force from 35N to 54N, in order to overcome the clamping force caused by contamination.

[0049] Although the invention has been described herein in conjunction with various embodiments, those skilled in the art will understand and implement other variations of the disclosed embodiments by reviewing the accompanying drawings, the disclosure, and the description of the drawings, in carrying out the claimed invention. In this specification, the word "comprising" does not exclude other components or steps, and "a" or "an" does not exclude multiple components. A single processor or other unit can implement several of the functions listed in the specification. While certain measures are described in different embodiments, this does not mean that these measures cannot be combined to produce good results.

[0050] Although the invention has been described in conjunction with specific features and embodiments, it is obvious that various modifications and combinations can be made therein without departing from the spirit and scope of the invention. Accordingly, this specification and drawings are merely illustrative of the invention and are considered to cover any and all modifications, variations, combinations, or equivalents within the scope of the invention. Clearly, those skilled in the art can make various alterations and modifications to the invention without departing from its spirit and scope. Thus, if such modifications and modifications fall within the scope of the invention and its equivalents, the invention is also intended to include such modifications and modifications.

Claims

1. A method for increasing the driving force of an electro-hydraulic servo valve based on a pre-stage, characterized in that, Includes the following steps: Configure a benchmark electro-hydraulic servo valve and a calibrated electro-hydraulic servo valve whose driving force needs to be increased; Obtain the benchmark pressure-current characteristic curve and the benchmark pressure-displacement characteristic curve of the front stage corresponding to the benchmark electro-hydraulic servo valve, and obtain the pressure-current characteristic curve and the pressure-displacement characteristic curve of the front stage corresponding to the electro-hydraulic servo valve to be calibrated. The benchmark displacement of the deflection plate of the benchmark electro-hydraulic servo valve under constant input current is obtained based on the benchmark pressure-current characteristic curve and the benchmark pressure-displacement characteristic curve of the pre-stage. The actual displacement of the deflection plate of the electro-hydraulic servo valve to be calibrated under constant input current is obtained based on the pressure-current characteristic curve and the pressure-displacement characteristic curve of the pre-stage. The theoretical displacement of the deflection plate of the electro-hydraulic servo valve to be calibrated under constant input current is calculated, and the pressure-displacement characteristic curve is further corrected by the deviation between the theoretical displacement and the actual displacement until the deviation between the actual displacement and the theoretical displacement is within the preset deviation range, so as to obtain the corrected pressure-displacement characteristic curve. Find the benchmark displacement of the benchmark electro-hydraulic servo valve under constant input current on the corrected pressure-displacement characteristic curve; The constraints are as follows: the open-loop transfer function gain of the electro-hydraulic servo valve from input current to output flow remains constant; the bandwidth variation of the electro-hydraulic servo valve is within a preset range. Based on the constraints, according to the preset driving force enhancement scheme, the theoretical displacement of the deflection plate of the electro-hydraulic servo valve to be calibrated under constant input current is increased to the benchmark displacement of the electro-hydraulic servo valve under constant input current. The preset driving force enhancement scheme includes: increasing the stiffness of the spring tube and the stiffness of the feedback rod, increasing the width of the nozzle jet orifice, and decreasing the width between the two receiving ports.

2. The method for increasing the driving force of an electro-hydraulic servo valve based on a pre-stage according to claim 1, characterized in that, The driving force was increased from 35N to 54N, the stiffness of the Bourdon tube was increased to 12.9Nm / rad, the stiffness of the feedback rod was increased to 1510 N / m, the nozzle jet width was increased to 0.163mm, and the width between the two receiving ports was reduced to 0.56mm.

3. The method for increasing the driving force of an electro-hydraulic servo valve based on a pre-stage according to claim 1, characterized in that, Develop a pre-defined driving force enhancement plan using the following methods: Construct a positive boosting model for the driving force of the electro-hydraulic servo valve to be calibrated: in, As the driving force; The effective area at both ends of the valve core; For preamplifier pressure gain; The distance from the rotation center of the armature baffle assembly to the center of the deflection plate; This refers to the electromagnetic torque coefficient of a permanent magnet torque motor. For the control current of the input control coil; The load torque generated by the fluid force of the pre-stage jet on the deflector plate is approximately zero; The moment of inertia of the armature assembly; For the complex field variable; is the viscous damping coefficient of the armature assembly; For the stiffness of the Bourdon tube; For the stiffness of the magnetic spring in a permanent magnet torque motor; For the feedback rod stiffness; The distance from the center of the deflection plate to the center of the feedback lever ball; The lifting parameters were determined based on the principle of not altering the overall structure of the armature assembly and the magnetic reluctance and magnetic flux of the torque motor stage. These lifting parameters included the stiffness of the Bourdon tube. Feedback rod stiffness Nozzle jet width and the width between the two receiving ports ; Configure constraints: in, Indicates the cross-sectional area of ​​the valve core; The net stiffness of the torque motor for the target electro-hydraulic servo valve; For the feedback rod stiffness; The distance from the center of the deflection plate to the center of the feedback lever ball; The distance from the rotation center of the armature baffle assembly to the center of the deflection plate; For pre-stage flow gain; The amplitude and bandwidth of the target electro-hydraulic servo valve; Minimum amplitude bandwidth for the target electro-hydraulic servo valve; For the target electro-hydraulic servo valve, the maximum amplitude bandwidth is required. is the open-loop amplification factor of the force feedback loop.

4. The method for increasing the driving force of an electro-hydraulic servo valve based on a pre-stage according to claim 3, characterized in that, Preamplifier pressure gain and nozzle jet width The following relationship must be satisfied: Preamplifier pressure gain and the width between the two receivers The following relationship must be satisfied: Pre-stage flow gain and nozzle jet width The following relationship must be satisfied: Preamplifier flow gain and the width between the two receiver ports The following relationship must be satisfied: In the formula, This indicates the pressure gain of the preamplifier stage. This indicates the pre-stage flow gain. Indicates the nozzle jet width. This indicates the width between the two receiving ports.

5. The method for increasing the driving force of an electro-hydraulic servo valve based on a pre-stage according to claim 1, characterized in that, The pressure-displacement characteristic curves of the benchmark electro-hydraulic servo valve and the pre-stage of the electro-hydraulic servo valve whose driving force needs to be increased were obtained through the following method: By setting different valve core displacements, the pressure difference between the two receiving ports corresponding to each displacement is obtained, and the pressure-displacement characteristic curve of the pre-stage is obtained by plotting points.

6. The method for increasing the driving force of an electro-hydraulic servo valve based on a pre-stage according to claim 1, characterized in that, The pressure-current characteristic curves of the benchmark electro-hydraulic servo valve and the electro-hydraulic servo valve whose driving force needs to be increased were obtained through the following method: By setting different input currents, the pressure difference between the two receiving ports corresponding to each input current is obtained, and the pressure-current characteristic curve is obtained by plotting points.

7. The method for increasing the driving force of an electro-hydraulic servo valve based on a pre-stage according to claim 1, characterized in that, The theoretical displacement is denoted as It is obtained by calculation as follows: The distance from the rotation center of the armature baffle assembly to the center of the deflection plate; For the armature corner.

8. The method for increasing the driving force of an electro-hydraulic servo valve based on a pre-stage according to claim 1, characterized in that, Armature corner It is obtained through the following calculation: For the overall stiffness of the torque motor; This refers to the mechanical damping ratio of the torque motor. For the complex field variable; The natural frequency of the armature baffle assembly; This refers to the electromagnetic torque coefficient of a permanent magnet torque motor. For the control current of the input control coil; The distance from the rotation center of the armature baffle assembly to the center of the deflection plate; The distance from the center of the deflection plate to the center of the feedback lever ball; For the feedback rod stiffness; For valve core displacement; The load torque generated by the fluid force of the pre-stage jet on the deflector plate is approximately zero.

9. An electro-hydraulic servo valve, characterized in that, The driving force of the electro-hydraulic servo valve based on the pre-stage, as described in any one of claims 1 to 8, is increased from 35N to 54N to overcome the clamping force caused by contamination.

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