JPH0333501A - Servo controller for fluid pressure actuator - Google Patents

Abstract

PURPOSE:To position a hydraulic actuator with high accuracy by calculating the deviation of a displacement command signal on the basis of the output of a load displacement detecting means and control valve displacement detecting means so as to compensate the displacement command signal in accordance with the deviation. CONSTITUTION:In a servo controller of an articulated mechanism such as a working robot arm, a displacement command drives a servo motor E through a characteristics compensator B and electric power amplifier, C to which a deviation controller A and a speed detector F return a signal, and further, a rotary control valve I and a rotary actuator K drives a load L via a deceleration gear G which receiving a mechanical feed-back. According to the detected displacement of the load L and to the detected rotation of the DC servo motor E, i.e., the detected displacement of the control valve, the deviation controller A detects a deviation of a displacement command signal 10 compensate the signal, thereby controlling the position of the rotary actuator K. Therefore, a highly accurate positioning and a stable control can be realized.

Description

Detailed Description of the Invention (Industrial Application Field) The present invention relates to a servo control device for a fluid pressure actuator, and more specifically, a DC servo motor drives a rotary valve to operate a vane motor to perform a required work. The present invention relates to a servo control device for a fluid pressure actuator that achieves highly accurate servo control without causing positional deviation or system oscillation.

(Prior art) One of the hydraulic switching valves is a rotary valve, and feed bank control is easy when the input/output is rotational motion, so it is often used to configure a servo control device. . For example, Figures 2 to 4, which will be described in detail later,
As shown in the figure, it is often seen that a rotary valve is inserted into the axis of an extraction swing motor and driven by a DC servo motor to form a servo actuator.

In such a rotary servo actuator, when the manual shaft is stationary, the servo valve (rotor valve) is closed, and the pressure in each oil chamber is oscillated. The pressure becomes the same, and the output shaft does not operate. If the input shaft is rotated by a DC servo motor for this assistance, one of the oil chambers will be connected to the high pressure ball l- of the servo valve, and the other will be connected to the low pressure boat, causing a difference in oil chamber pressure and causing rocking. The motor rotates. Next, when the input shaft is stopped, the output shaft rotates to the same angle as the human power shaft, the valve closes, each oil chamber has the same pressure, and the swing motor stops. In this case, by arranging the rotary valve so that the direction of rotation of the human power shaft matches the direction of rotation of the output shaft, the vane motor will rotate to the same extent in the direction that the rotary valve rotates, and there will be a gap between the valve and the output shaft. Mechanical position feedback occurs, thus allowing load positioning control by controlling the valve position.

An example of conventional servo control using this is shown in FIG.

In this conventional technology, a displacement detector is inserted between a servo motor and a rotary type control valve (rotary valve),
Position control is performed by feeding back the amount of valve displacement. Further, as described above, the displacement of the load is configured to mechanically follow the position of the valve by mechanical feedback.

(Problem to be Solved by the Invention) However, in this prior art, the mechanical feedback loop generates a force commensurate with the load when a certain load such as gravity is applied, which causes positional deviation between the valve and the actuator. Therefore, there is a problem that a deviation occurs between the command position and the load position, making it impossible to perform positioning control with high accuracy.

To explain this point with reference to FIG. 12, this figure is a block diagram that specifically rewrites the control method conceptually shown in FIG.
The valve angle θV determined based on 00 m is detected and fed back through angle detection means such as a rotary encoder attached to a servo motor. In this servo system, the output angle θL of the actuator matches the valve angle θV in the no-load state, but when a disturbance load τd occurs, there is no integral element before the input of τd (θV and θ1. (because the deviation of θ is not integrated through the integral element), there is an inconvenience that a deviation occurs between the manual angle θν and the output angle θL (in addition, in the same figure, k
P: Position loop gain, kν: Velocity loop, Ts
: Delay time, S nira plus operator, a, b: Actuator parameters, A: Actuator cross-sectional area, JL
: load and actuator inertia). In order to solve this problem, it is conceivable to increase the supply pressure to reduce the deviation, that is, to increase the parameter a, but in that case, a new problem arises regarding the pressure resistance of the piping, accumulator, etc.

Furthermore, in the prior art, a method has been proposed in which the position of the load is fed back instead of the position of the valve. Although it is possible to eliminate deviations using this method, it also creates the problem that the system oscillates due to interference with mechanical feedback. FIG. 13 is a block diagram showing the conventional method. In this example, a detection means such as a potentiometer is used to detect and feed back the load position θL. In this case, there is an integral element "1/s" in front of the human power of the disturbance load τd, and the deviation becomes zero. However, in the illustrated example, there are two integral elements ``1/s, 1/A s'' in the -circuit loop, which causes a phase delay of 180 degrees.There is a phase advance element in the feedback loop. However, in any case, the feedback system is quite unstable, and even if the calculation at the first addition point is performed using a microcomputer, the calculation takes time, and the phase becomes even more unstable. In addition, mechanical feedbanks, as is obvious from their construction, contain nonlinear elements such as dead zones (generally the boat does not open when the valve moves), which causes the system to oscillate. As will be explained with reference to the experimental data in FIG. 15, oscillation occurs when an experiment is performed.

Therefore, an object of the present invention is to eliminate the above-mentioned drawbacks of the prior art, and to provide a servo control device for a fluid pressure actuator that enables highly accurate positioning control without causing positional deviation or system oscillation. There is a particular thing.

(Means for Solving the Problems) In order to achieve the above-mentioned objects, the present invention provides a driving means that operates in response to a displacement command signal, and a fluid pressure control valve that is connected to the driving means and is displaced in accordance with the operation thereof. , a fluid pressure actuator that is connected to the fluid control valve, operates to follow the displacement of the fluid pressure actuator, and supplies fluid pressure energy to the connected load; Control valve displacement amount detection means for detecting the displacement amount of the fluid pressure control valve, and the output of the control valve displacement amount detection means are inputted to calculate the deviation from the displacement command signal, and a displacement command signal is generated according to the deviation. In the servo control device for a fluid pressure actuator, the servo control device for a fluid pressure actuator is provided with a load displacement detection means for detecting the displacement amount of the load, and the correction means includes the load displacement amount detection means and the control valve displacement amount detection means. The deviation from the displacement command signal is calculated by inputting the output of the displacement command signal, and the displacement command signal is corrected according to the deviation.

(Operation) FIG. 1 is an explanatory diagram conceptually showing a servo control device according to the present invention. As shown in the figure, the difference from the conventional method is that the amount of displacement of the valve and the amount of displacement of the load are detected at the same time. The controller compares the detected displacement amounts of the pulp and load with the displacement command value, calculates and outputs the manipulated variable. This makes it possible to realize a servo control system with higher precision than conventional techniques.

(Example) Hereinafter, a servo control device for a fluid pressure hydrator according to the present invention will be explained with reference to the accompanying drawings. First, its structure will be explained with reference to FIGS. 2 to 5.

FIG. 2 is an explanatory sectional view generally showing the servo control device according to the present invention, which is disposed at one of the joints of a multi-joint mechanism such as a working robot arm. In the figure, the lower arm 1
The upper end of 0 is connected to a casing 11 that houses a hydraulic swing (vane) motor M, and an output shaft 13 of the swing motor is rotatable within the space formed by the casing 11 and cover 12. It is bearing. A spline 14 is carved on the right end side of the output shaft 13 in the same figure.
The output torque of the swing motor is transmitted via this spline 14 to the other arm 20 located above.

The lower end of the upper arm 20 has a yoke shape, and in the figure, a yoke 21 on the left side is rotatably connected to the cover 12 by a bearing 22, and oil passages 24 and 25 are formed inside the yoke 23 on the right side. At the same time, the right yoke 23 is connected to the output shaft 13 via a joint 26 so as to hug the output shaft. The joint 26 is fixed to the yoke 23 via a plurality of bolts 27 while meshing with the spline 14 in order to transmit the above-mentioned torque. Yoke 21 and yoke 23 are yoke bodies 21a, 23a and cap 21b.
, 23b, each of which is fixed with a bolt. Further, a washer 28 for fixing the bearing 22 is attached to the yoke 21 with a countersunk screw 29.

On the left side of FIG. 2, a motor flange 15 for regulating the movement of the bearing 22 in the axial direction is also attached to the cover 12 via bolts 16. The motor flange 15 is molded integrally with the casing 101 of the DC servo motor 100, as will be described later. In addition, in order to detect the amount of displacement of the load described above through the movement angle of the swing motor, a potentiometer 1 is used.
04 is provided at an appropriate position on the arm 20 side, and a timing heald 17 for transmitting the movement angle to the potentiometer 104 and a helt pulley 1 for hanging the heel.
8 105 is provided.

The output shaft 13 of the swing motor M described above is rotatably supported by a second bearing 30 provided on the casing 11 and a third bearing 31 provided on the cover 12. A pressure oil supply oil passage 32 is formed in the axial center of the output shaft 13, and a return oil passage 33 is bored in parallel thereto. At the portion connected to the yoke 23 described above, these two oil passages 32 and 33 are connected to the oil passage 2 through a suitable sealing material 34.
It is connected to 4.25. Further, a sleeve 35 constituting a rotary valve is press-fitted into the left half of the output shaft 13, and two annular grooves 3 surround this sleeve 35.
6.37 is provided.

The portion of the output shaft 13 sandwiched between the second and third bearings 3031 has a relatively large diameter and forms a rotating piston portion 40 of the swing motor M.
The cover 12 and the donut-shaped side plate 4 are sandwiched between the cover 12 and the donut-shaped side plate 4.
2 is incorporated. The shape of the rotating piston part 40 is the third
As clearly shown in the figure, it is housed eccentrically in a cylinder 52, and its upper part is fan-shaped and has a small hole 44 provided parallel to the output shaft 13, and a small hole 44 extending outward in the radial direction starting from this small hole. A pair of grooves 46° and 47 having a constant width are bored. Three vanes 4 are slidably inserted into these grooves 46 and 47.
8a and 48b are inserted. The first groove 46 is inserted into one of the working chambers 50 defined by the rotary piston part 40 via the oil passages 45 and 38 inside the rotary piston part 40 and the groove 41 drilled in the side surface thereof. The other groove 47 is also a groove 49.39 inside the rotary piston part 40 and an excavated groove 4.
3 to a further working chamber 51 . The six vanes 48a (48b) are forced outward by the hydraulic pressure generated in each small hole 44, and inwardly by the hydraulic pressure acting on their tips, and the resultant force is applied to some extent by the vane 48a.
(48b) acts to push it outward, but is held in the illustrated position in contact with the cylinder inner surface 52a by the cylinder 52 formed around the working chamber. Reference numeral 64 indicates a relief oil passage for preventing pumping.

A stator 55 is fixed at a lower position within the cylinder 52,
The stator 55 has the same length in the axial direction as the cylinder 52, and similarly has vertical grooves 57 and 58, in which a plurality of second moons 59a (59b) are slidably housed. The second vane 59 a (59 b) is completely the same in structure and function as the piston side vane 43 a (48 b) described above, except that the second vane 59 a (59 b) can be slid inward in the radial direction. The hydraulic pressure of each working chamber 50 (51) is connected to the small hole 56 of each vertical groove 57, 58 that accommodates the second hem.
0.61. The stator 55 is a swing motor M
When the piston operates, it receives a reaction force from the piston, so as shown in FIG.
2 protrudes and is fixed 3 is inserted into the hole drilled on the stator side, and the side plate 42
A second pin 63 having a relatively small diameter is also provided between the stator 55 and the side plate 42 in order to prevent the rotor from rotating due to the rotation of the swing motor. Hydraulic pressure is supplied to the right side of the side plate 42, and serves to constantly press the side plate 42 toward the cylinder 52 to keep the side clearance of the swing motor constant. As shown in FIG. 3, since the center of rotation of the rotary piston part 40 is eccentric in the cylinder 52, the rotational center of the rotary piston part 40 and the cylinder inner surface 5 are
2a increases as the vane 48a (48b) swings counterclockwise, and the degree of protrusion of the vane 48a (48b) also increases. kept within range.

Reference numeral 80 indicates a hydraulic rotary valve, which cooperates with the aforementioned sleeve 35 to constitute a rotation guide valve. This rotary valve 80 is well known, and is rotatably housed within the sleeve 35.
Based on the relative rotational change with respect to position 5, one of the working chambers 50 and 51 shown in FIG. It drives and controls the
Formally, it is a type of four-way switching valve. rotary valve 8
0 is housed in a chamber formed in the swing motor output shaft 13 with its right end side in contact with a wall surface, and is disposed at the illustrated position where its left end side is blocked by a thrust bearing 92. Further, the low-return valve 80 is bored at an axial position corresponding to the pressure oil supply passage 32 bored at the axial position of the output shaft 13, and an oil passage 65 is formed therein. The oil passage 6
A pipe 96 is bridged to connect 5 with the pressure oil supply path 32.

As clearly shown in FIG. 4 (cross-sectional view taken along the line IV-IV in FIG. 3), the oil passage 65 formed at the pulp axis position is 18
It is connected to three grooves 67a among six grooves carved in the axial direction on the cylindrical surface through three oil passages 66 that branch at 0 degree intervals. Further, the remaining three grooves 67b communicate with the return oil passage 33 bored in the output shaft 13. On the other hand, sleeve 35
Six windows 82 and 83 are bored in the interior to cooperate with the windows 82 and 83, and the windows 82 and 83 are connected to the working chamber 50 of the swing motor via the pressure oil supply line 38, and the window 83 is connected to the The two working chambers 51 are connected to the other working chamber 51 through return oil passages 39 . A throttle 97 of an appropriate diameter is protruded inside the pipe 96, which determines the maximum amount of oil supplied to the swing motor M, and also creates a minute thrust force in the differential pressure generated there. The rotary valve 80 is pressed against the thrust bearing 92 and fixed in the illustrated position.

A DC servo motor 100 is provided 3 to provide input rotation to the rotary encoder 80, and its output shaft 84
A pin 86 extending in the radial direction protrudes from and is smoothly inserted into a groove 87 prepared on the low revalve 80 side. A harmonic reducer 106 is provided adjacent to the DC servo motor 100, and the motor output is reduced by an appropriate factor to amplify the torque, which is then output from the output shaft 84. A rotary encoder 102 is provided at the other end of the harmonic reducer 106 to detect the displacement amount of the rotary valve 80 through the rotation angle of the DC servo motor 100.

In addition, in order to supply compressed citron to the swing motor M or return it to the tank side, an oil passage consisting of two pipes 2, 3, 4.5 is provided inside the arms 10 and 20. The pressure well supplied through the casing 11 is guided to the oil passage 32 in the shaft center of the output shaft 13 described above through the hole 6 bored in the casing 11, and is supplied to the high pressure port of the low-pressure valve 80.

In addition, the return hydraulic oil is supplied through an oil passage 33 provided in the output shaft 13.
and is led to the pipe 3 described above through a hole 8 in the casing. On the arm 20 side, a pipe 4 for supplying to the next joint (not shown) and a pipe 5 for the return oil phase are prepared. In FIG. 1, reference numerals 6a and 24a (25a) indicate plugs that close the open ends of the oil passages.

When the DC servo motor 100 receives a command from a control unit to be described later and rotates the output shaft 84, the rotor valve 80 is also relatively displaced to switch between the two and one of the working chambers 50 and 51 as a hydraulic source. and connect the other end to the tank. That is, in FIG. 4, when the rotary valve 80 is turned clockwise, for example, the window 82 is connected to the high pressure side, the window 83 is connected to the tank side, and the rotary piston part 4
A pressure difference is generated on the left and right sides of the oscillation motor, and the swing motor is hydraulically driven, and its output shaft 13 rotates clockwise. Next, the relative relationship between the rotary valve 80 and the sleeve 35 returns to the original state and becomes a neutral state, but in this way, the movement of the rotary valve 80 and the output shaft 13 of the swing motor are rotated in synchronization due to mechanical feedback. It is constructed so that a system exists. The rotational force of the output shaft 13 is transmitted through the spline 14 to
A bending moment is generated between the two arms 10, 20, and the two arms articulate with respect to each other around the rocking motor.

As shown, a hydraulic pump 112 driven by an electric motor 110 is provided to pump up hydraulic oil from a tank, and the pressurized hydraulic oil is sent to the pipe 2 via a high pressure line 114. Eight accumulators 116 are connected to the high pressure line 114. A control unit 12 for controlling the operation of the electric motor 110 and DC servo motor 100 described above.
0 is set. The control unit 120 includes the rotary encoder 102 and the potentiometer 104 described above, as well as the oil pressure sensor 12 inserted at an appropriate position in the high pressure line.
2 outputs are sent out.

FIG. 5 is a block diagram showing details of the control unit 120. As shown in the figure, the control unit 120 includes a CPU 120
a, ROM 120b, RAMI 20c, bus 120d, A/D conversion circuit 12120i 20f, counter board 122g, and D/A conversion circuit 12120i 20i. D conversion circuit 12120i and RAMI20c
is stored primarily in . Further, the output signal θen of the rotary encoder 102 is input to the microcomputer via the counter board 120g, and is also input to the motor drive circuit 126 via the F/V conversion circuit 124. In the microcomputer, the CPU 120a outputs the angle command value θcom, calculates the deviation from the detected angle θpotθen, and outputs the angle command value θcom, as will be described later.
120i and output to the motor drive circuit 126 as a speed command value. In the motor drive circuit 126, the pulse output of the low-resolution encoder 102 is converted into a speed value by the F/V conversion circuit 124, and the DC servo motor 100 follows the command speed.
control so that it rotates.

Further, the output of the oil pressure sensor 122 installed in the high pressure line 114 is input to the microcomputer through an A/D conversion circuit 12120f, and the CPU 120a calculates the deviation between the detected oil pressure and the target oil pressure, and converts it to the D/A conversion circuit. 121
201 and outputs a speed command value, and a pump system J consisting of an electric motor 110 and a hydraulic pump 112.
is controlled so that the pressure of the accumulator 116 is kept constant.Next, the operation of this control device will be explained with reference to the flow chart of FIG.

First, the CPU 120a in the SIO searches the ROM 120b and reads the angle command value θcom set for the time shown in FIG. 7 (the ?lJJ value of the integral element S is zero).
Next, in S12, the valve angle θen is read from the output of the rotary encoder 102, and in S14, the angle θen of the output shaft 13 on the load side is read from the output of the potentiometer 104.
After reading the pot, the integral element S is calculated as follows in S16.

S=S+ (θcom-θpot) Subsequently, in 31B, the speed command value VC is calculated as follows.

VC=kP(θcom-θen+kPΔθ(θc.

m-θpot)+kl Δθ・S) Here, kP, kPΔθ: Proportional gain klΔθ: Integral gain S: Laplace operator Next, the speed command value VC calculated in S20 is outputted as 1, and the command angle and actual angle are output in 322. The loop is repeated until the difference between the values decreases below a predetermined value (set as appropriate).

Regarding the above, please refer to the block diagram in Figure 8. When T, the deviation (-θVθL) is not integrated in this control device either. However, in this control device, as shown in the figure, the actual load angle θpot and the angle command value θcow
The deviation from l is proportionally integrated and added to the valve command value Vc. That is, the actual load angle θpot and the angle command value θc
By proportionally integrating the deviation with om, the deviation is made zero, and the actual valve angle θen is also fed back to improve the instability of the prior art in which only the load position is fed back. With such a configuration, the deviation can be effectively reduced to zero without causing system oscillation.

FIG. 9 shows experimental data showing the responsiveness of step input in this control device. In addition, Fig. 14 shows experimental data showing the responsiveness of the conventional method that detects only the valve angle shown in Fig. 12, and Fig. 14 shows the responsiveness of the conventional method that detects only the load angle shown in Fig. 13. This is the experimental data shown. As shown in the figure, a deviation remains with the method of detecting only the valve angle, and oscillations occur with the method of detecting only the load angle, but it can be seen that in the device according to the present invention, control is stable and almost no deviation occurs. Furthermore, Fig. 10 shows experimental data when a step-like disturbance torque is applied to the device according to the present invention, and even in such a state, the disturbance converges within a short time and the output shaft angle is well within the target value. It will be understood that they are following.

In this embodiment, the rotary encoder 102
The valve angle is detected through the rotation angle of the sarpo motor 100, and the angle of the output shaft 13 on the load side is also detected through the potentiometer 104, and the angle command value is corrected based on these detected values. Alternatively, stable and highly accurate servo control can be realized without causing system oscillation. Furthermore, since the angle of the output shaft is also detected, there is no need to consider gear backlash, etc., and it is sufficient to attach the rotary encoder 102 to the shaft end of the DC servo motor at the front stage of the reducer 106. It is easy to attach and the configuration of the device is simplified.

In the above embodiments, the present invention has been explained using a swing motor as an example of a fluid pressure actuator, but the present invention is not limited to this, and a hydraulic cylinder may also be used. good.

(Effects of the Invention) The servo control device for a fluid pressure actuator according to claim 1 is provided with a load displacement amount detection means for detecting the displacement amount of the load, and the correction means is configured to detect the load displacement amount detection means and the control valve displacement amount. The deviation from the displacement command signal is calculated by inputting the output of the means, and the displacement command signal is corrected according to the deviation, so that no deviation occurs between the command value and the displacement value of the fluid pressure actuator. It is possible to perform highly accurate positioning without any interference, and to achieve stable control by eliminating interference with mechanical feedback.

In the servo control device for a fluid pressure actuator according to claim 2, the drive means includes a reduction gear that reduces the output thereof, and the control valve displacement amount detection means is provided upstream of the reduction gear. This makes it possible to provide the detection means for detecting the amount of displacement of the control valve at the front stage of the reduction gear of the drive means, which also has the secondary effect of simplifying the structure.

In other words, in the conventional technology, the detection is performed at the rear stage of the reducer in consideration of gear backlash, etc., and the detection means such as the rotary encoder or tachogenerator in the motor is installed on the non-output side. Since this is common, it is difficult to install a detection means when trying to detect it at the latter stage of the reducer.However, in the present invention, since the displacement on the load side is detected, backlash of the gears of the reducer can no longer be taken into account. The detection means can be attached to the front stage of the reduction gear and detect the displacement amount of the control valve through the shaft position of the drive means, which facilitates the installation work and simplifies the structure.

[Brief explanation of drawings]

5. Fig. 1 is an explanatory diagram conceptually showing a servo control device for a fluid pressure actuator according to the present invention, Fig. 2 is an explanatory cross-sectional view showing the device disposed on the joint axis of a working robot arm, and Fig. 3 is an explanatory diagram conceptually showing the servo control device for a fluid pressure actuator according to the present invention. Figure 2 is a sectional view along line ■-■, Figure 4 is Figure 3 rV-I
5 is an explanatory block diagram showing details of the control unit in FIG. 2, FIG. 6 is an explanatory flow chart showing the operation of the device according to the present invention, and FIG. 7 is the flow chart. 8 is a block diagram specifically showing the control method according to the present invention shown in FIG. 1, and FIG. 9 is an explanatory diagram showing the angle command value data stored in the control unit used in the present invention. Experimental data showing the stability of control of the device related to
Similarly, Figure 11 shows experimental data showing the stability of control when a step disturbance torque is applied to the device according to the present invention.
The figure is an explanatory diagram conceptually showing the conventional technology in comparison with Fig. 1, Fig. 12 is a block diagram more specifically showing the conventional art shown in Fig. 11, and Fig. 13 is a concrete diagram showing another conventional art. 14 is the experimental data showing the stability of the control of the prior art shown in FIGS. 11 and 12, and FIG. 15 is the experimental data showing the stability of the control of the prior art shown in FIG. 13. The data is out. M... Swing motor (fluid pressure actuator), 2.3
, 4.5... Pipe, 68... Gauging hole, 10
．． 20...Arm, 11...Casing, 12...
・Cover, 13...ti Moving motor output shaft, 14...
・Spline, 15...Motor flange, 18, 10
5...Belt pulley, 21.23...Yoke, 24
．． 25...Oil passage, 26...Joint, 22.30 3
DESCRIPTION OF SYMBOLS 1... Bearing, 32... Pressure oil supply oil path, 33... Return oil path, 34... Seal material, 35... Stub IJ-bu,
36.37... annular groove, 38.39... groove, 40.
...Rotating piston part, 4143...Groove, 42...Side plate, 44.56...Small hole, 45.49...Oil passage, 4
6.47...Groove, 48a, 48b, 59a, 59b.
・ ・Höhn, 50.51... Swing motor operating chamber,
52... Cylinder, 55... Stator, 57.5B,
60°61...Groove, 62.63...Pin, 6465
．． 66.68...Oil passage, 67a, 67b...Groove (boat), 80...Lower valve (Lower type control valve L) 82, 83...Window, 84°...DC motor output shaft, 86 ...Pin, 87...Groove, 92...Thrust bearing, 93...Oil passage, 96...Pipe, 97...
・Aperture, 100...DC servo motor (drive means),
101...DC motor casing, 102...Rotary encoder (control valve displacement detection means), 104...
・Potentiometer (load displacement detection means L 106
...Harmonic reducer, 110...Electric motor,
112...Hydraulic pump, 114...High pressure line, 1
16...Accumulator, 120...Control unit (correction means), 120a-...CPU, 120b...
ROM, 120c--RAM, 120d-bus, 12
0e, 12Of-...A/D conversion circuit, 120g...
Counter board, 120h, i... D/A conversion circuit, 122... Oil pressure sensor, 124... F/V conversion circuit, 126... Motor drive circuit, 128... Pump system

Claims (2)

[Claims] (1) a. A driving means that operates in response to a displacement command signal; b. A fluid pressure control valve that is connected to the driving means and is displaced in response to its operation. c. a fluid pressure actuator that operates in accordance with the fluid pressure control valve and supplies fluid pressure energy to a connected load; control valve displacement amount detection means; and (e) correction means for inputting the output of the control valve displacement amount detection means, calculating a deviation from the displacement command signal, and correcting the displacement command signal according to the deviation. A servo control device for a fluid pressure actuator, further comprising: f, load displacement detection means for detecting the displacement amount of the load, and the correction means receives outputs from the load displacement amount detection means and the control valve displacement amount detection means. A servo control device for a fluid pressure actuator, characterized in that the deviation from the displacement command signal is calculated, and the displacement command signal is corrected according to the deviation. (2) The fluid pressure actuator according to claim 1, wherein the drive means includes a reduction gear that reduces the output thereof, and the control valve displacement amount detection means is provided at a stage upstream of the reduction gear. servo control device.