CN104730935A - Oneway-friction loading type electro-hydraulic load simulator without surplus torque - Google Patents

Abstract

A one-way friction loading type electro-hydraulic load simulator without redundant torque belongs to the field of electro-hydraulic servo control and automatic control. The invention solves the problem of redundant torque that seriously affects the loading performance existing in the loading process of the existing electro-hydraulic load simulator. In the loading system of the present invention, the hydraulic cylinder driven by the electro-hydraulic flow servo valve applies a given force to the friction pair composed of a pair of friction discs that maintain constant relative rotation. Through the relative rotation between the friction discs, the applied force It is converted into corresponding torque, which is loaded to the loaded object through the transmission device, and at the same time, the torque is collected by the torque sensor and fed back to the real-time control industrial computer of the system, so as to realize the precise closed-loop control of the system. This electro-hydraulic load simulator has no redundant torque, high loading accuracy, high system frequency band, and simple and reliable control algorithm.

Description

One-way friction-loaded electro-hydraulic load simulator without excess torque

technical field

The invention relates to a semi-physical simulation model, which belongs to the field of electro-hydraulic servo control and automatic control.

Background technique

In the production of national defense and military industries such as aerospace, nautical ships, and weaponry, as well as civil industrial production such as construction engineering, earthquake engineering, automotive engineering, bioengineering, and agricultural machinery engineering, it is usually necessary to test the dynamic load performance of the product to ensure the design. Product performance. On the one hand, most dynamic loads are arbitrary forces (torques) that vary with space and time, and have strong non-controllability, such as air hinge dynamic moments, seawater wave forces, etc.; Manpower and material resources, some are even impossible to realize, such as seismic fluctuation load. These reasons lead to and promote the emergence and development of ground hardware-in-the-loop simulation technology. Ground hardware-in-the-loop simulation technology is to reproduce various factors and parameters in the working process of the measured object under laboratory conditions, and transform the classic self-destructive experiment into prediction research under laboratory conditions. Ground hardware-in-the-loop simulation technology has the advantages of good controllability, non-destructive, all-weather and simple and convenient operation, and this kind of experiment is repeatable, and its economy is unmatched by classical self-destructive experiments. The load simulator is a device used to simulate the force (moment) load required by the object under laboratory conditions. It is mainly used to simulate the aerodynamic moment load on the rudder surface of the aircraft during flight. Spectrum.

There have been many technical problems in the development of traditional electro-hydraulic load simulators: (1) The existence of redundant torque seriously affects the control performance of the loading system. The steering gear system and the loading system of the electro-hydraulic load simulator are approximately rigidly connected together. When the steering gear actively moves, it will inevitably generate strong disturbances to the loading system, causing excess torque. The value is related to the motion state of the steering gear, and it is more serious when starting and reversing. (2) It is difficult to realize the accuracy of dynamic loading. The electro-hydraulic load simulator requires the loading system to simulate the aerodynamic torque experienced by the aircraft during flight. Generally, the torque signal is an arbitrary function. To accurately reproduce this function, the loading system is required to be a high-order static-free system. However, the existence of excess torque, especially its differential characteristics, makes it difficult to achieve high-order static error in the loading system, especially when the steering gear system has a high motion frequency. (3) It is difficult to guarantee the system performance under small moment loading. Due to the influence of factors such as servo valve dead zone and pressure fluctuation, it is difficult to guarantee the system performance when the electro-hydraulic load simulator is loaded with small torque. In addition, when loading with small torque, the impact of excess torque on the loading system will become relatively significant, which will reduce the accuracy of the loading system, make it difficult to guarantee the loading sensitivity, and even submerge the loading signal, making the system unable to achieve normal loading. (4) The loading system is complex to control. The actual loaded object of the electro-hydraulic load simulator is the steering gear of various aircraft, and the structural parameters of the steering gear are related to the performance of the loading system. Different types of steering gear may lead to changes in the control performance of the system, especially the changes in the parameters of the redundant torque compensation control link, so the control system is required to have a certain degree of robustness. This makes the control of the loading system more complex and difficult.

Contents of the invention

The purpose of the present invention is to solve the problem of excess torque that seriously affects the loading performance existing in the loading process of the existing electro-hydraulic load simulator, and provides a one-way friction loading type electro-hydraulic load simulator without excess torque.

In order to completely eliminate the redundant moment of the load simulator, improve the accuracy of dynamic loading, and achieve accurate small moment loading while obtaining simple structure, low cost and simple control strategy, a friction loading method is proposed, which does not exist by The excess torque generated by the main motion of the tested steering gear seriously interferes with the loading performance, so that the loading performance can be comprehensively improved, accurate dynamic loading can be achieved, small torque loading performance can be guaranteed, and the complexity of the control strategy can be simplified. The proposal of this kind of friction-loaded electro-hydraulic load simulator without redundant torque complies with the trend of increasing the loading performance of the load simulator with the improvement of the maneuverability and control accuracy of aircraft and missiles, which promotes the advancement of the national defense industry and can bring great benefits. good economy.

The one-way friction loading type electro-hydraulic load simulator without redundant torque of the present invention includes an industrial computer, an A/D data acquisition card, a servo amplifier, a DSP motion control card, a torque motor, a code disc, a large servo valve, and a torque sensor , D/A conversion circuit and load simulation unit;

The code disc and the large servo valve are set on the steering gear under test; the signal output end of the code disc is connected with the first input end of the A/D data acquisition card, and the signal output end of the torque sensor is connected with the second input end of the A/D data acquisition card The output end of the A/D data acquisition card is connected to the signal feedback end of the industrial computer; the command output end of the industrial computer is connected to the input end of the D/A conversion circuit, and the output end of the D/A conversion circuit is connected to the servo amplifier The input end of the servo amplifier is connected to the input end of the servo amplifier, and the output end of the given angular displacement command of the servo amplifier is connected to the command input end of the large servo valve; the output end of the given torque signal of the servo amplifier is connected to the command input end of the servo valve;

The load simulation unit includes friction disc B, friction disc A, base, large gear, transmission shaft, pinion, sliding disc, thrust bearing, feather key, transition disc, force sensor, three sets of springs, and three hydraulic cylinder covers , three hydraulic cylinders, servo valve, displacement sensor, valve block, main shaft and second base;

The industrial computer controls the rotation of the torque motor through the DSP motion control card. The torque motor drives the pinion to rotate in a constant direction through the transmission shaft. The steering gear is rigidly connected to the main shaft through a torque sensor, and the large gear loads torque on the measured steering gear through the main shaft;

The friction plate A is fixed on one side of the large gear, the friction plate B is set on the side of the friction plate A away from the large gear, and the friction plate B is fixed on the sliding plate, and the sliding plate is connected to the main shaft through a sliding key;

The servo valve is set on the valve block, the valve block is set on the second base, and the second base is set on the outer surface of the main shaft. The second base is disc-shaped and evenly distributed along the outer circumference of the second base. Three rectangular grooves, the opening of the rectangular groove faces the large gear and the outer circle; each rectangular groove is provided with a hydraulic cylinder cover facing the outer circle opening, and the second base is provided with a transition plate facing the large gear, and the transition plate covers Hold the axial opening of each rectangular groove, and the transition plate is frictionally connected with the hydraulic cylinder cover plate in the axial direction; a hydraulic cylinder, a displacement sensor and a spring are arranged in each rectangular groove, and the displacement sensor, hydraulic cylinder and spring They are connected axially in sequence and pressed against the transition plate, and the piston of the hydraulic cylinder is pressed against the spring; the axial space between the transition plate and the sliding plate is provided with a thrust bearing, and a force sensor is embedded in the transition plate, and the force sensor is clipped Between the thrust bearing and the transition plate; the piston displacement signal of the hydraulic cylinder collected by the displacement sensor is fed back to the industrial computer, and the force signal generated by the hydraulic cylinder collected by the force sensor is fed back to the industrial computer.

The advantages of the present invention: the friction loading type electro-hydraulic load simulator without redundant torque proposed by the present invention has no redundant torque generated by the main motion of the steering gear to be tested that seriously interferes with the loading performance, and has a compact structure. Since there is no redundant moment, when the steering gear performs any form of main motion, compared with the traditional structure of the electro-hydraulic load simulator, this kind of friction-loaded electro-hydraulic load simulator can obtain dynamic loading with higher precision, and can achieve high Accurate small-amplitude moment loading can simplify the complexity of the control strategy, the system control strategy is more general and reliable, and the loading system has the characteristics of high precision, high dynamics, and high frequency response. The proposal of this kind of friction-loaded electro-hydraulic load simulator without redundant torque can comprehensively improve the accuracy of torque loading and realize the active loading of torque. One loading system can be used for loading under different loading conditions without redesigning the corresponding The controller enables the load simulator to be more easily and widely used without re-adjustment by professional technicians.

Description of drawings

Fig. 1 is the functional block diagram of the one-way friction loading type no redundant moment electro-hydraulic load simulator of the present invention;

Fig. 2 is a torque closed-loop control block diagram;

Fig. 3 is a block diagram of angular displacement closed-loop control;

Fig. 4 is a structural schematic diagram of a load simulation unit;

Fig. 5 is a cross-sectional view along line A-A of Fig. 4 .

Detailed ways

Specific embodiment one: the present embodiment is described below in conjunction with Fig. 1 to Fig. 5, and the unidirectional friction loading type electro-hydraulic load simulator without redundant moment described in the present embodiment, it comprises industrial computer 101, A/D data acquisition card 102, Servo amplifier 103, DSP motion control card 104, torque motor 105, code disc 106, large servo valve 107, torque sensor 108, D/A conversion circuit 109 and load simulation unit;

The code disc 106 and the large servo valve 107 are arranged on the steering gear under test; the signal output end of the code disc 106 is connected with the first input end of the A/D data acquisition card 102, and the signal output end of the torque sensor 108 is connected with the A/D data acquisition The second input end of card 102 is connected, and the output end of A/D data acquisition card 102 is connected with the signal feedback end of industrial computer 101; The command output end of industrial computer 101 is connected with the input end of D/A conversion circuit 109, D/ The output end of the A conversion circuit 109 is connected with the input end of the servo amplifier 103, and the given angular displacement command output end of the servo amplifier 103 is connected with the large servo valve 107 command input end; the given torque signal output end of the servo amplifier 103 is connected with the servo valve 25 command input end is connected;

The load simulation unit includes B friction plate 1, A friction plate 2, base 10, large gear 11, transmission shaft 12, pinion 16, sliding plate 17, thrust bearing 18, feather key 19, transition plate 20, force sensor 21 , three groups of springs 22, three hydraulic cylinder cover plates 23, three hydraulic cylinders 24, servo valve 25, displacement sensor 26, valve block 27, main shaft 28 and second base 31;

The industrial computer 101 controls the rotation of the torque motor 105 through the DSP motion control card 104, and the torque motor 105 drives the pinion 16 to rotate in a constant direction through the transmission shaft 12, and the pinion 16 drives the large gear 11 with a constant speed through the meshing form, and with the pinion For relative steering rotation, the steering gear under test is rigidly connected with the main shaft 28 through the torque sensor 108, and the large gear 11 loads the torque on the steering gear under test through the main shaft 28;

The A friction disc 2 is fixed on one side of the large gear 11, the B friction disc 1 is set on the side of the A friction disc 2 away from the large gear 11, and the B friction disc 1 is fixed on the sliding disc 17, and the sliding disc 17 and the main shaft 28 pass through The feather key 19 is connected; in this way, the B friction disc 1 can slide axially based on the main shaft 28 and can transmit torque to the main shaft 28 at the same time. Thereby ensuring that friction disc B 1 can transmit the axial pressure exerted by hydraulic cylinder 24 without loss, ensuring that the pressure between friction disc A 2 and friction disc B 1 is consistent with the pressure exerted by hydraulic cylinder 24, and the sliding key connection can connect friction disc A 2. The friction torque generated between the B friction discs 1 is transmitted to the main shaft 28, so as to finally act on the steering gear under test.

The servo valve 25 is arranged on the valve block 27, the valve block is arranged on the second base 31, and the outer circular surface of the main shaft 28 is provided with the second base 31, the second base 31 is disc-shaped, along the second base Three rectangular grooves are evenly distributed on the outer circumference of seat 31, and the opening of rectangular groove faces bull gear 11 and outer circle; The gear 11 side is provided with a transition plate 20, the transition plate 20 covers the axial opening of each rectangular groove, and the transition plate 20 is frictionally connected with the hydraulic cylinder cover plate 23 in the axial direction; a hydraulic cylinder is arranged in each rectangular groove 24. A displacement sensor 26 and a spring 22, the displacement sensor 26, the hydraulic cylinder 24 and the spring 22 are axially connected in turn, and pressed against the transition plate 20, and the piston of the hydraulic cylinder 24 is pressed against the spring 22; the transition plate 20 and the spring 22 The axial space between the sliding discs 17 is provided with a thrust bearing 18, and the transition disc 20 is embedded with a force sensor 21, and the force sensor 21 is sandwiched between the thrust bearing 18 and the transition disc 20; the hydraulic cylinder 24 collected by the displacement sensor 26 The piston displacement signal is fed back to the industrial computer 101, and the force signal generated by the hydraulic cylinder 24 collected by the force sensor 21 is fed back to the industrial computer 101.

A friction disc 2 is fixed on the side of the bull gear 11 by riveting and gluing, so that its rotational speed and direction of rotation are consistent with the bull gear 11 . B friction disc 1 and A friction disc 2 constitute a friction pair.

The torque motor 105 is connected with the transmission shaft 12 through a coupling. The torque motor 105 is used to drive the A friction disc 2 to rotate, so as to ensure that the rotational speed of the torque motor 105 can be controlled in real time, so that the A friction disc 2 and the B friction disc 1 can obtain the optimal relative speed under different loading conditions, ensuring the torque loading type The frictional heat generated between A friction disc 2 and B friction disc 1 is the smallest, and the friction impact between A friction disc 2 and B friction disc 1 is suppressed, so that the moment loading is more stable.

The torque motor 105 drives the B friction disc 1 to rotate at a speed higher than the maximum speed of the main movement of the tested steering gear, so that when the tested steering gear moves in any form, the B friction disc 1 and A friction disc 2 maintain a constant direction relative rotation.

The transition between the B friction disc 1 and the three circumferentially evenly distributed hydraulic cylinders 24 is through the thrust bearing 18, so as to ensure that when the B friction disc 1 connected with the main shaft 28 through the feather key performs the main movement together with the steering gear under test, the swinging B friction The friction torque between disc 1 and the static hydraulic cylinder 24 fixed on the base 10 is very small and negligible. When the hydraulic cylinder 24 applies axial pressure between A friction disc 2 and B friction disc 1, it will not be affected by the steering gear under test. Main motion interference, B friction disc 1 will not produce other interference torque with other components except friction torque with A friction disc 2.

Three hydraulic cylinders 24 evenly distributed in the circumferential direction apply axial pressure to B friction disc 1 under the drive of electro-hydraulic flow servo valve 25, so as to ensure that the applied pressure can be evenly distributed on A friction disc 2 and B friction disc 1 The contact surface makes the torque loading more stable and easier for servo control. In addition, when the three hydraulic cylinders 24 are loaded at the same time, they can obtain a smaller volume under the condition of obtaining the same output force, so as to ensure that the system can obtain the maximum loading torque and at the same time It can ensure that the system has a sufficient bandwidth.

The industrial computer 101 is also equipped with a set of real-time control software. The industrial computer 101 controls the torque motor 105 through the DSP motion control card 104 to drive the pinion 16 to rotate in a constant direction through the transmission shaft 12 at an optimal speed; The rotating speed of this rotating speed is always higher than the maximum rotating speed of the steering gear under test, and rotates by the relative direction of rotation with pinion 16.

The industrial computer 101 sends an instruction to the servo valve 25, and the servo valve 25 drives the three hydraulic cylinders 24 to work, so that it applies force to the B friction disc 1 evenly according to the given control signal, and the applied force passes through the spring 22 and the transition plate 20 , force sensor 21, thrust bearing 18, and sliding disc 17 act on B friction disc 1; the piston displacement of hydraulic cylinder 24 is measured by displacement sensor 26 and fed back to industrial computer 101; the force generated by hydraulic cylinder 24 is measured and fed back by force sensor 21 Give the industrial computer 101. Assuming that the friction coefficient between A friction disc 2 and B friction disc 1 is constant, according to the generation principle of friction force:

f=F·μ

In the formula: f - friction force (N);

F——The pressure (N) exerted by the hydraulic cylinder 24 between A friction disc 2 and B friction disc 1;

μ——The coefficient of friction between A friction disc 2 and B friction disc 1.

The contact area between A friction disc 2 and B friction disc 1 is circular, and it can be considered that the force F applied by the hydraulic cylinder 24 to the A friction disc 2 and B friction disc 1 and the corresponding friction force f are evenly distributed in the On the area of the circular ring, it is easy to draw from the knowledge of calculus that the torque T acting on the main shaft 28 through the conversion of the friction force f through the friction disc is:

$\mathrm{<span\; class="notranslate">\; T</span>}<span\; class="notranslate">\; =</span>{<span\; class="notranslate">\; \&Integral;</span>}_{\mathrm{<span\; class="notranslate">\; a</span>}<span\; class="notranslate">\; /</span>\mathrm{<span\; class="notranslate">\; 2</span>}}^{\mathrm{<span\; class="notranslate">\; b</span>}<span\; class="notranslate">\; /</span>\mathrm{<span\; class="notranslate">\; 2</span>}}\frac{{\mathrm{<span\; class="notranslate">\; 8</span>}\mathrm{<span\; class="notranslate">\; r</span>}}^{\mathrm{<span\; class="notranslate">\; 2</span>}}\mathrm{<span\; class="notranslate">\; f</span>}}{{\mathrm{<span\; class="notranslate">\; b</span>}}^{\mathrm{<span\; class="notranslate">\; 2</span>}}<span\; class="notranslate">\; -</span>{\mathrm{<span\; class="notranslate">\; a</span>}}^{\mathrm{<span\; class="notranslate">\; 2</span>}}}\mathrm{<span\; class="notranslate">\; dr</span>}$

In the formula: f - friction force (N);

b——Outer diameter of the contact ring between A friction disc 2 and B friction disc 1 (m);

a——Inner diameter of contact ring between A friction disc 2 and B friction disc 1 (m).

It can be seen from the above that since there is friction between the friction disc B 1 and the friction disc A 2, and the rotation speed of the friction disc A 2 is always higher than the main movement speed of the steering gear under test, so the friction disc B 1 applies the control signal to the friction disc A 2 The resulting pressure F is transformed into a corresponding torque T through the friction and relative rotation between A friction disc 2 and B friction disc 1 . Since the B friction disc 1 is connected to the main shaft 28 through the sliding key 19 through the sliding disc 17, the transition between the B friction disc 1 and the hydraulic cylinder 24 is through the thrust bearing 18 and the rotation speed of the A friction disc 2 is higher than the maximum rotation speed of the steering gear. Disc 2 and B friction disc 1 will always keep relative rotation in a constant direction, so that the main motion of the steering gear under test will not interfere with hydraulic cylinder 24 exerting pressure on B friction disc 1, and the main motion of the steering gear will not affect the resulting Torque T produces interference, that is, the load simulator does not have excess torque generated by the main motion of the servo under test. The torque T generated between the A friction disc 2 and the B friction disc 1 will be transmitted to the steering gear under test through the connection of the feather key 19 as mentioned above and the main shaft 28, so as to realize the unidirectional torque loading on the steering gear. The generated torque T is measured by a torque sensor and fed back to the industrial computer 101 through the A/D data acquisition card 102 . The real-time control software will use the torque signal fed back by the given expected torque signal to calculate the control signal according to the designed controller, and transmit the calculated control signal to the servo valve 25 through the D/A conversion circuit 109 and the servo amplifier 103 Apply pressure between the A friction disc 2 and the B friction disc 1 by driving the hydraulic cylinder 24 . Furthermore, a torque T is generated under the relative rotation of the A friction disc 2 and the B friction disc 1, thus forming a torque loading closed-loop system, and its control block diagram is shown in Fig. 2 . The steering gear system under test is generally closed-loop servo control. As shown in Figure 1, the angular displacement of the analog steering gear system here is measured by the code disc and fed back to the industrial computer 101 through the A/D data acquisition card 102, and the real-time control software will use the given expected angular displacement signal and the feedback The angular displacement signal calculates the control signal according to the designed industrial computer 101 and transmits it to the large servo valve 107 through the D/A conversion circuit 109 and the servo amplifier 103 to drive the swing hydraulic motor steering gear to rotate, thus forming the analog steering gear system Angular displacement closed-loop control, its control block diagram is shown in Figure 3.

The displacement of the piston of the hydraulic cylinder 24 and the thrust generated by the hydraulic cylinder 24 can be measured in real time by the displacement sensor 26 and the force sensor 21 respectively, and the displacement of the hydraulic cylinder 24, the thrust of the hydraulic rod and the torque are fed back to the industrial computer 101 through the A/D data acquisition card 102 In this way, the friction performance between A friction disc 2 and B friction disc 1 can be tested and analyzed by using piston displacement, thrust, torque, etc. Stability, etc., so that the friction-loaded electro-hydraulic load simulator can be used as a dynamic performance testing machine for friction materials.

The friction-loaded electro-hydraulic load simulator described in this embodiment is used to accurately simulate the actual moment load spectrum of the load-carrying object under laboratory conditions. When the liquid load simulator is loaded, it is seriously disturbed by the redundant torque generated by the main motion of the loaded object, which realizes high-precision, high-frequency response, and high dynamic moment loading of the loaded object under any main motion of the loaded object, reducing the The complexity of the loading control algorithm of the load simulator completely solves the influence of excess torque on a series of destructive loading performances of the load simulator. In the loading system of the present invention, the hydraulic cylinder driven by the electro-hydraulic flow servo valve applies a given force to the friction pair composed of a pair of friction discs that maintain constant relative rotation. Through the relative rotation between the friction discs, the applied force The force is converted into the corresponding torque, which is loaded on the loaded object through the transmission device, and the torque is collected by the torque sensor and fed back to the real-time controller of the system, so as to realize the precise closed-loop control of the system. This electro-hydraulic load simulator has the advantages of no redundant torque, high loading accuracy, high system frequency band, simple and reliable control algorithm, and excellent small-amplitude loading performance. Simulation can provide economical, high-precision, and high-reliability equipment support for the performance improvement of the loaded steering gear system.

Specific Embodiment 2: This embodiment will further describe Embodiment 1. The large gear 11 and the main shaft 28 are connected through the first angular contact ball bearing 3 and the second angular contact ball bearing 4, and are connected by the stop washer 7 and the round nut 9. The bull gear 11 is fixed axially on the main shaft 28 .

Specific Embodiment 3: This embodiment further explains Embodiment 1. The main shaft 28 and the transmission shaft 12 are fixed together by the first base 10, and the first base 10 and the second base 31 are respectively located on both sides of the large gear 11. , the outer circumference of the transmission shaft 12 between the first base 10 and the pinion 16 is provided with a sleeve 15; the first base 10 is fixed to the main shaft 28 through the first end cover 6, several bolts 5 and the third contact ball bearing 8 together; the first base 10 is fixed together with the pinion 16 through the second end cover 13 , several bolts 5 and the fourth contact ball bearing 14 .

Embodiment 4: In this embodiment, Embodiment 1 is further described. A third base 32 is provided on the main shaft 28 on the same side as the first base 10 . The third base 32 is located at the end of the main shaft 28 .

The configuration of this embodiment is used to connect with external components to meet rigid requirements.

Embodiment 5: This embodiment will further explain Embodiment 1. The valve block 27 realizes that the servo valve 25 communicates with the oil inlet 29 and the oil outlet 30. The servo valve 25 is also provided with an A port and a B port. The B ports are respectively connected with the oil inlets of the three hydraulic cylinders 24; the B ports are respectively connected with the oil outlets of the three hydraulic cylinders 24.

Port A and port B of the servo valve 25 communicate with port A and port B of the hydraulic cylinder 24 through pipelines arranged in the second base 31 (port A and port B represent the other two ports or ports of the servo valve 25 respectively). Two oil ports of the hydraulic cylinder 24).

In this embodiment, the pipeline in the valve block 27 and the second base 31 realizes the connection with the external oil circuit through an oil inlet 29 and an oil outlet 30, and supplies and discharges oil to three hydraulic cylinders 24 at the same time .

Claims (5)

1. unidirectional friction-loaded formula is without Surplus Moment electrohydraulic load simulator, it is characterized in that, it comprises industrial computer (101), A/D data collecting card (102), servoamplifier (103), DSP motion control card (104), torque motor (105), code-disc (106), large servo-valve (107), torque sensor (108), D/A change-over circuit (109) and load simulation unit;

Code-disc (106) and large servo-valve (107) are arranged on tested steering wheel; The signal output part of code-disc (106) is connected with the first input end of A/D data collecting card (102), the signal output part of torque sensor (108) is connected with the second input end of A/D data collecting card (102), and the output terminal of A/D data collecting card (102) is connected with the signal feedback end of industrial computer (101); The instruction output end of industrial computer (101) is connected with the input end of D/A change-over circuit (109), the output terminal of D/A change-over circuit (109) is connected with the input end of servoamplifier (103), and the given angular displacement instruction output end of servoamplifier (103) is connected with large servo-valve (107) command input; The given moment signal output part of servoamplifier (103) is connected with the command input of servo-valve (25);

Load simulation unit comprises B frictional disk (1), A frictional disk (2), pedestal (10), gear wheel (11), transmission shaft (12), pinion wheel (16), slider disc (17), thrust bearing (18), feather key (19), transition disc (20), force snesor (21), three groups of springs (22), three hydraulic cylinder cover plates (23), three hydraulic cylinders (24), servo-valve (25), displacement transducer (26), valve block (27), main shaft (28) and the second pedestal (31),

Industrial computer (101) is rotated by DSP motion control card (104) control moment motor (105), torque motor (105) is rotated by constant direction by transmission shaft (12) driving pinion (16), pinion wheel (16) drives gear wheel (11) with constant rotating speed by mesh form and relative with pinion wheel turns to rotation, tested steering wheel is rigidly connected by torque sensor (108) and main shaft (28), gear wheel (11) by main shaft (28) to tested Loading for actuator moment;

A frictional disk (2) is fixed on a side of gear wheel (11), B frictional disk (1) is arranged on the side that A frictional disk (2) deviates from gear wheel (11), and B frictional disk (1) is fixed in slider disc (17), slider disc (17) is connected by feather key (19) with main shaft (28);

Servo-valve (25) is arranged on valve block (27), valve block (27) is arranged on the second pedestal (31), the outer round surface of main shaft (28) is provided with the second pedestal (31), second pedestal (31) is disc, uniform three rectangular recess of excircle along the second pedestal (31), the opening surface of rectangular recess is to gear wheel (11) and cylindrical; Each rectangular recess arranges a hydraulic cylinder cover plate (23) towards cylindrical opening, second pedestal (31) is provided with transition disc (20) towards gear wheel (11) side, transition disc (20) covers the axially open of each rectangular recess, transition disc (20) vertically with hydraulic cylinder cover plate (23) frictional connection; A hydraulic cylinder (24), a displacement transducer (26) and a spring (22) are set in each rectangular recess, displacement transducer (26), hydraulic cylinder (24) and spring (22) are axially connected successively, and top is pressed on transition disc (20), the piston top of hydraulic cylinder (24) is on spring (22); Axial space between transition disc (20) and slider disc (17) is provided with thrust bearing (18), transition disc (20) is embedded with force snesor (21), and force snesor (21) is clipped between thrust bearing (18) and transition disc (20); The piston displacement signal of the hydraulic cylinder (24) that displacement transducer (26) gathers feeds back to industrial computer (101), and the force signal that the hydraulic cylinder (24) that force snesor (21) gathers produces feeds back to industrial computer (101).

2. according to claim 1 unidirectional friction-loaded formula without Surplus Moment electrohydraulic load simulator, it is characterized in that, gear wheel (11) is connected by the first angular contact ball bearing (3) and the second angular contact ball bearing (4) with main shaft (28), and is axially fixed on main shaft (28) by gear wheel (11) by stop washer (7) and round nut (9).

3. according to claim 1 unidirectional friction-loaded formula without Surplus Moment electrohydraulic load simulator, it is characterized in that, main shaft (28) and transmission shaft (12) adopt the first pedestal (10) to be fixed together, first pedestal (10) and the second pedestal (31) lay respectively at the both sides of gear wheel (11), and transmission shaft (12) excircle between the first pedestal (10) and pinion wheel (16) is provided with sleeve (15); First pedestal (10) is fixed together by the first end cap (6), some bolts (5) and the 3rd contact ball bearing (8) and main shaft (28); First pedestal (10) is fixed together by the second end cap (13), some bolts (5) and the 4th contact ball bearing (14) and pinion wheel (16).

4. according to claim 1 unidirectional friction-loaded formula without Surplus Moment electrohydraulic load simulator, it is characterized in that, with the main shaft (28) of the first pedestal (10) homonymy on be also provided with the 3rd pedestal (32), the 3rd pedestal (32) is positioned at the end of main shaft (28).

5. according to claim 1 unidirectional friction-loaded formula without Surplus Moment electrohydraulic load simulator, it is characterized in that, valve block (27) realizes servo-valve (25) and is connected with oil-out 30 with oil-in 29, servo-valve (25) is also provided with A mouth and B mouth, and described A mouth is connected with the oil-in of three hydraulic cylinders (24) respectively; Described B mouth is connected with the oil-out of three hydraulic cylinders (24) respectively.