JP2021156351A - Hydraulic drive position controller - Google Patents

Abstract

To provide a hydraulic drive position controller capable of shortening a time from start of control for controlling a positioning target to start of actual movement.SOLUTION: A hydraulic actuator 11 generates a thrust for moving a positioning target. Hydraulic oil supply systems 21, 22 supply hydraulic oil to the hydraulic actuator. A controller 50 supplies the hydraulic oil to the hydraulic actuator from the hydraulic oil supply systems for each control cycle. During the period from a stationary state to the start of movement of the positioning target, the controller makes a supply amount of hydraulic oil in a first control cycle larger than a supply amount of hydraulic oil in second and subsequent control cycles.SELECTED DRAWING: Figure 3

Description

The present invention relates to a control device for a hydraulic drive position control system that controls the supply of hydraulic oil to a hydraulic actuator that moves a positioning target to a command position.

Patent Document 1 below discloses a hydraulic drive position control system capable of supplying hydraulic oil to a hydraulic cylinder to improve the accuracy of minute displacement of the piston position. In the hydraulic drive position control system disclosed in Patent Document 1, hydraulic oil is intermittently supplied to the hydraulic cylinder to adjust the piston position to the target position.

Japanese Unexamined Patent Publication No. 2018-173131

In a system that intermittently supplies hydraulic oil to a hydraulic cylinder at each control cycle and moves it step by step to a target position, the position specified by the position command changes stepwise due to the influence of static friction force, and then the hydraulic pressure The positioning target fixed to the cylinder does not start moving immediately, causing wasted time and deteriorating the positioning response. This is because the hydraulic oil needs to be supplied by the control output a plurality of times until the differential pressure of the hydraulic cylinder reaches the maximum static friction equivalent pressure (hereinafter referred to as friction equivalent pressure). The number of control outputs required for the positioning target to start moving × the time of the control cycle is a wasted time that does not contribute to the movement. When the positioning target starts to move, the differential pressure and the friction equivalent pressure are balanced, so the hydraulic oil supplied as the control output contributes to the displacement. Further, when the positioning target starts to move, the friction state changes from static friction to dynamic friction and the load pressure drops, so that the displacement becomes easier. Wasted time may occur when the direction of movement of the positioning target changes. When the direction of motion is reversed, the direction of static friction force is also reversed, so the balance between the differential pressure and the friction equivalent pressure is lost, and until the differential pressure reaches the friction equivalent pressure, the number of control outputs x the wasted time of the control cycle occurs. Deteriorate the positioning response.

An object of the present invention is to provide a control device capable of reducing the wasted time of a positioning response caused by the influence of static friction acting on a positioning target in a position control system driven by a flood control.

According to one aspect of the invention
A control device that drives a hydraulic actuator by controlling a hydraulic oil supply system that supplies hydraulic oil to a hydraulic actuator that generates thrust to move a positioning object.
The hydraulic oil is supplied from the hydraulic oil supply system to the hydraulic actuator at each control cycle.
A control device is provided in which the supply amount of hydraulic oil in the first control cycle is larger than the supply amount of hydraulic oil in the second and subsequent control cycles during the period from the stationary state to the start of movement of the positioning target.

By increasing the supply amount of hydraulic oil in the first control cycle, it is possible to shorten the time until the positioning target in the stationary state starts to move.

FIG. 1 is a schematic view of a hydraulic drive position control system equipped with a control device according to an embodiment. FIG. 2A is a cross-sectional view of a sliding surface between a positioning target of a hydraulic drive position control system equipped with a control device according to an embodiment and a support member supporting the positioning target, and FIG. 2B is a cross-sectional view of a sliding surface when the sliding surface is moved. It is a graph which shows the generated friction force and friction coefficient. FIG. 3A is a block diagram of the control system of the hydraulic drive position control system according to the present embodiment, and FIG. 3B is a graph showing an example of the time waveform of the position (command position) Xref commanded by the position command. FIG. 4 shows the command position Xref commanded by the position command of the hydraulic drive position control system according to the comparative example, the operation pulse output by the pulse generating unit (FIG. 3A), the differential pressure P of the hydraulic cylinder, and the position X of the positioning target. It is a graph which shows the time waveform of the position deviation e. FIG. 5 shows the command position Xref when the position of the positioning target is controlled by using the hydraulic drive device according to the embodiment, the operation pulse output by the pulse generating unit (FIG. 3A), the differential pressure P of the hydraulic cylinder, and the position of the positioning target. It is a graph which shows the time change of X, the position deviation e.

The hydraulic drive position control system according to the embodiment will be described with reference to FIGS. 1 to 5.
FIG. 1 is a schematic view of a hydraulic drive position control system according to an embodiment. Both rod type hydraulic cylinders 11 include a cylinder body 11S, a piston 11P, a first rod 11A, and a second rod 11B. The first rod 11A, the second rod 11B, and the piston 11P are fixed, and the cylinder body 11S moves with respect to the fixed portion. The positioning target 10 is attached to the cylinder body 11S. The positioning target 10 is, for example, a moving table of a grinding device.

FIG. 2A is a cross-sectional view of the positioning target 10 and the support member 18 that supports the positioning target 10. A hydraulic cylinder 11 is attached to the lower surface of the positioning target 10. The support member 18 movably supports the positioning target 10 in the moving direction of the cylinder body 11S.

Two rows of convex guide sliding surfaces (convex) 10A having an inverted triangular cross section are provided on the lower surface of the positioning target 10. The positioning target 10 is supported by a guide sliding surface (concave) 18A having a concave V-section provided on the support member 18, and the hydraulic cylinder 11 can be moved in the axial direction. An oil film of lubricating oil is formed between the guide sliding surface (convex) 10A and the guide sliding surface (concave) 18A to be fitted. When the positioning target 10 is driven from the stationary state, as shown in FIG. 2B, a large static friction force acts on the sliding surface as compared with the moving state, which deteriorates the positioning response and the positioning accuracy.

In the cylinder body 11S shown in FIG. 1, a first oil chamber 13A on the first rod 11A side and a second oil chamber 13B on the second rod 11B side separated by the piston 11P are provided. A first port 14A and a second port 14B are provided at the ends of the first rod 11A and the second rod 11B, respectively. The first port 14A is connected to the first oil chamber 13A via a first oil passage 12A provided in the first rod 11A. The second port 14B is connected to the second oil chamber 13B via a second oil passage 12B provided in the second rod 11B.

The hydraulic oil supply system 20 supplies the hydraulic oil to the hydraulic cylinder 11 and recovers the hydraulic oil from the hydraulic cylinder 11. The hydraulic oil supply system 20 includes a hydraulic pump 21 and an electric motor 22. The hydraulic pump 21 is a bidirectional hydraulic pump, and is driven by an electric motor 22 in both forward and reverse directions.

A relief valve 23A is connected to an oil passage between the hydraulic pump 21 and the first port 14A. Similarly, the relief valve 23B is connected to the oil passage between the hydraulic pump 21 and the second port 14B. The relief valves 23A and 23B release the hydraulic oil in the oil passage to the hydraulic oil tank 30 when the pressure in the oil passage becomes equal to or higher than a predetermined pressure.

The charge pump 26, which is a one-way hydraulic pump, is driven by the electric motor 27. When the pressure in the oil passage between the first port 14A and the hydraulic pump 21 and the pressure in the oil passage between the second port 14B and the hydraulic pump 21 become lower than the predetermined charge pressure, the pressure in the oil passage becomes lower than the predetermined charge pressure, via the shuttle valve 24. The hydraulic oil is introduced into the oil passage from the charge pump 26. When the pressure in the oil passage on the discharge side of the charge pump 26 becomes equal to or higher than a predetermined pressure, the relief valve 28 releases the hydraulic oil in the oil passage to the hydraulic oil tank 30.

The pressure sensors 15A and 15B measure the pressure in the oil passage connected to the first port 14A and the second port 14B, respectively. The measured value of pressure is input to the control device 50. The difference between the measured values of the pressure sensors 15A and 15B, that is, the product of the differential pressure and the pressure receiving area of the piston 11P is the thrust generated by the hydraulic cylinder 11. That is, the two pressure sensors 15A and 15B function as thrust sensors for measuring the physical quantity related to the thrust generated by the hydraulic cylinder 11. The position sensor 16 measures the position of the cylinder body 11S, that is, the positioning target 10. The measured position is input to the control device 50.

The control device 50 controls the hydraulic oil supply system 20 based on the pressure and position measured values input from the pressure sensors 15A and 15B, and the position sensor 16, and moves the positioning target 10 to the commanded position. Information for designating the command position is externally given to the control device 50 as a position command.

Next, a method of controlling the thrust applied to the positioning target 10 by the hydraulic cylinder 11 will be described.

The thrust F for moving the positioning target 10 is expressed by the following equation using the pressure receiving area A of the piston 11P and the differential pressure P between the first oil chamber 13A and the second oil chamber 13B. The differential pressure P is a value obtained by dividing the thrust F of the hydraulic cylinder 11 by the pressure receiving area A of the hydraulic cylinder. Since the thrust F is generated as a reaction of the load resistance such as the frictional force and the inertial resistance, the differential pressure P is also called the load pressure.

The differential pressure P between the first oil chamber 13A and the second oil chamber 13B is the volume elastic coefficient Kv of the hydraulic oil and the volume of the hydraulic oil discharged from the hydraulic pump 21 to the hydraulic cylinder 11 from the definition of compressibility of the hydraulic oil. It is expressed by the following equation using ΔV, the hydraulic cylinder 11, and the volume V _{0 of the hydraulic oil pipeline connected to the hydraulic cylinder 11.}

The volume ΔV of the hydraulic oil discharged from the hydraulic pump 21 to the hydraulic cylinder 11 is as follows, using the push-out volume Dp per rotation of the hydraulic pump 21 and the rotation angle displacement Δθ [rad] of the input shaft of the hydraulic pump 21. It is represented by an expression.

The rotation angle displacement Δθ of the input shaft of the hydraulic pump 21 is expressed by the following equation using the angular velocity ω of the input shaft and the drive time Δt of the electric motor 22. The angular velocity ω is a preset fixed value.

From the formulas (1) to (4), the differential pressure P of the hydraulic cylinder 11 is represented by the following formula (5), and the generated thrust F is represented by the following formula (6).

Equation (6) shows that the thrust F can be changed by controlling the drive time Δt of the electric motor 22 and the hydraulic pump 21. In this embodiment, the hydraulic pump 21 is driven by a pulse time width (hereinafter, referred to as “pulse width”) shorter than the control cycle for each predetermined control cycle, and the hydraulic pump 21 is stopped for the remaining time. As a result, the positioning target 10 is moved. That is, the positioning target 10 is moved to the target position by executing the control cycle a plurality of times while intermittently supplying hydraulic oil to the hydraulic cylinder 11 every control cycle to finely feed the positioning target 10. In the present specification, such control is referred to as "discharge capacity control".

The discharge capacity control is adopted in a region where the positional deviation of the positioning target 10 is minute, and as a result, the moving speed of the positioning target 10 at the time of positioning is minute. For example, it is used in a minute velocity region where a stick-slip phenomenon occurs.

On the other hand, discharge flow rate control is adopted in a region where the position deviation or speed of the positioning target 10 is large, for example, in a speed region where the stick-slip phenomenon does not occur. In the discharge flow rate control, the rotation speed of the hydraulic pump 21 is controlled according to the position deviation of the positioning target 10.

Since speed fluctuations are likely to occur in a very low speed region where a stick-slip phenomenon occurs, it is difficult to stably control the speed by controlling the discharge flow rate. Therefore, it is difficult to stably align the position of the positioning target 10 with the target position. By adopting the discharge capacity control instead of the discharge flow rate control in the low speed region, the position of the positioning target 10 can be accurately aligned with the target position.

Next, with reference to FIGS. 3A and 3B, position control to which the discharge capacity control of the positioning target 10 by the hydraulic drive position control system according to the present embodiment is applied will be described.

FIG. 3A is a block diagram of the control system of the hydraulic drive position control system according to the present embodiment. The control device 50 gives an operation pulse to the motor driver 25 every control cycle. The motor driver 25 supplies driving power to the electric motor 22 based on the information commanded by the operation pulse. The hydraulic pump 21 supplies hydraulic oil to the hydraulic cylinder 11 by the power given from the electric motor 22. The hydraulic cylinder 11 generates thrust by the hydraulic oil supplied from the hydraulic pump 21 to move the positioning target 10 (FIG. 1). As a result, the position X of the positioning target 10 changes. The position sensor 16 (FIG. 1) measures the position X of the positioning target 10.

The control device 50 includes a deviation / pulse width conversion unit 51, a frictional force compensation unit 52, and a pulse generation unit 53. The command position Xref of the positioning target 10 is input to the control device 50. The control device 50 calculates the position deviation e by subtracting the position X from the command position Xref.

The deviation / pulse width conversion unit 51 generates the first pulse width Tw1 as the first manipulated variable according to the position deviation e. For example, the first pulse width Tw1 is proportional to the position deviation e. The first pulse width Tw1 may be made constant when the position deviation e exceeds a predetermined value.

The measured value of the differential pressure P between the first oil chamber 13A and the second oil chamber 13B of the hydraulic cylinder 11 (FIG. 1) is input to the friction force compensating unit 52 of the control device 50. The differential pressure P is calculated by subtracting the measured value of the other pressure sensor 15B from the measured value of one pressure sensor 15A. This differential pressure P is a physical quantity corresponding to the thrust generated by the hydraulic cylinder 11. The frictional force compensating unit 52 generates a second pulse width Tw2 as a second manipulated variable based on the position deviation e and the differential pressure P. The detailed processing of the frictional force compensating unit 52 will be described later.

The control device 50 adds the second pulse width Tw2 to the first pulse width Tw1 to generate the third pulse width Tw3 as the third manipulated variable. The third pulse width Tw3 is input to the pulse generation unit 53. The pulse generation unit 53 gives an operation pulse to the motor driver 25 based on a preset fixed rotation speed set value ω and a third pulse width Tw3. The pulse width of the operation pulse is equal to the third pulse width Tw3, and the wave height of the pulse is equal to the rotation speed set value ω. As described above, the operation pulse includes information on the time for driving the electric motor 22 and information on the rotation speed of the electric motor 22.

The motor driver 25 drives the electric motor 22 at a rotation speed equal to the rotation speed set value ω for a time equal to the third pulse width Tw3.

FIG. 3B is a graph showing an example of the time change of the command position Xref given to the control device 50. The horizontal axis represents time and the vertical axis represents command position Xref. The cycle for changing the command position Xref (hereinafter referred to as the step cycle) is referred to as Tstp, and the amount of change in the command position Xref (hereinafter referred to as the step amount) is referred to as ΔXstp.

Next, the processing of the frictional force compensating unit 52 will be described.
A frictional resistance proportional to the weight acts on the positioning target 10, and the positioning target 10 does not move unless a force equal to or greater than the frictional force is applied. The maximum static friction force Ffrc is the frictional force at the moment when the vehicle starts to move from the stationary state. When the thrust F of the hydraulic cylinder 11 exceeds the maximum static friction force acting on the positioning target 10, the positioning target 10 starts to move. The thrust F of the hydraulic cylinder 11 is the product of the pressure difference (differential pressure P) between the first oil chamber 13A and the second oil chamber 13B partitioned by the piston 11P and the pressure receiving area A. The value obtained by dividing the maximum static friction force Ffrc by the pressure receiving area A is referred to as the friction equivalent pressure Pfrc. That is, the friction equivalent pressure Pfrc is obtained by converting the maximum static friction force Ffrc acting on the positioning target 10 into the load pressure of the hydraulic cylinder 11. When the differential pressure P of the hydraulic cylinder 11 exceeds the friction equivalent pressure Pfrc, the positioning target 10 starts to move. On the other hand, the differential pressure P of the hydraulic cylinder 11 can be adjusted by the drive time Δt of the hydraulic pump 21 as shown in the equation (5).

The frictional force compensating unit 52 generates the second pulse width Tw2 using the following equation (7) based on the position deviation e and the differential pressure P. Here, it is assumed that the differential pressure P of the hydraulic cylinder 11 is positive and the positioning target 10 moves in the positive direction.

Here, Tfrc is a constant of proportionality. Tfrc is the pulse width required to generate the friction equivalent pressure Pfrc from the state where the differential pressure P is zero. sgn (e) indicates a sign function, which is +1 when the position deviation e is positive (e> 0) and -1 when the position deviation e is negative (e <0). The second pulse width Tw2 is proportional to the difference between the friction equivalent pressure Pfrc and the differential pressure P and the position deviation e. The pressure deviation Pfrc-P is divided by Pfrc, and the position deviation e is divided by ΔXstp for normalization. When the position deviation e becomes the maximum ΔXstp and the differential pressure P is zero at the rising (or falling) timing of the step in which the command position Xref changes, Tw2 is a proportional constant as shown in the following equation (8). Consistent with Tfrc.

In the second pulse width Tw2 of the equation (7), as shown in the equation (8), Tw2 becomes the maximum value, that is, Tfrc at the rising timing of the command position Xref. Further, under the condition that the position deviation e is constant, the second pulse width Tw2 is proportional to the difference between the friction equivalent pressure Pfrc and the differential pressure P of the hydraulic cylinder 11.

In the equation (7), the sign function sgn (e) is added to the friction equivalent pressure Pfrc because the maximum static friction force acts in the direction opposite to the movement direction, so that the correction is made according to the movement direction. This is because it is necessary to consider the sign of Pfrc. For example, considering the case where the step amount ΔXstp is switched from positive to negative, the position deviation e becomes negative, so that the sign function sgn (e) becomes -1. The differential pressure P at this time is positive (because the pressure in the + direction before inversion remains). At this time, the equation (7) can be modified as in the following equation (9), and the second pulse width Tw2 becomes larger than Tfrc. When the differential pressure P works in the direction opposite to the direction of the static friction force in this way, the second pulse width Tw2 is wider than Tfrc, and the differential pressure P is increased in the first pulse after the step amount ΔXstp is switched from positive to negative. The pressure equivalent to friction is reached.

Next, the physical significance of the equation (7) will be described.
Since the differential pressure P is smaller than the friction equivalent pressure Pfrc during the period from the stationary state to the start of movement of the positioning target 10, even if an operation pulse is given to the motor driver 25 to supply hydraulic oil to the hydraulic cylinder 11, the cylinder of the hydraulic cylinder 11 The body 11S cannot be displaced, the supplied hydraulic oil is used for pressurization, and the differential pressure P increases each time a pulse is applied. That is, the differential pressure P is the lowest in the case of the first control cycle immediately after the start time of the step cycle Tstp (FIG. 3B) (that is, immediately after the command position Xref changes in a step shape), and the number of control cycles is the lowest. It increases as it increases, and the difference between the friction equivalent pressure Pfrc and the differential pressure P becomes smaller. As shown by the equation (7), the second pulse width Tw2 also increases in proportion to the difference between the friction equivalent pressure Pfrc and the differential pressure P, so that the second pulse width of the first control cycle having the lowest differential pressure P Tw2 is the largest. Therefore, in the first control cycle, more hydraulic oil can be supplied as compared with the second and subsequent control cycles. As a result, the rise of the differential pressure P can be accelerated, and as a result, the number of control cycles until the differential pressure P reaches the friction equivalent pressure Pfrc can be reduced. When the differential pressure P exceeds the friction equivalent pressure Pfrc, the positioning target 10 starts to move, and frictional force compensation becomes unnecessary. At this time, the friction equivalent pressure Pfrc and the differential pressure P are balanced, and the difference between the two approaches zero, so that the second pulse width Tw2 also becomes a small value, and the effect of friction compensation decreases.

Further, in the equation (7), the second pulse width Tw2 is also proportional to the position deviation e. The intention is described below. The period in which friction compensation is required is the period from the start of the step cycle Tstp (FIG. 3B) to the start of movement of the positioning target 10. The position deviation e in this period is equal to the step amount ΔXstp (FIG. 3B). The term e / ΔXstp relating to the position deviation of the equation (7) is 1, and the second pulse width Tw2 is determined only by the relationship between the friction equivalent pressure Pfrc and the differential pressure P. When the positioning target 10 starts to move toward the command position Xref, the position deviation e decreases. The term related to the position deviation at this time is less than 1. For example, when the position deviation e is 10% of the step amount ΔXstp, the term related to the position deviation becomes 0.1. Since this value is multiplied by the pressure term, the effect of the pressure term is diminished by a factor of 10. This means that the effect of frictional force compensation weakens as the positioning target 10 approaches the command position Xrefni.

Next, the excellent effects of the above-described embodiment will be described with reference to FIGS. 4 and 5.

FIG. 4 shows the command position Xref commanded by the position command of the hydraulic drive position control system according to the comparative example, the operation pulse generated by the pulse generating unit (FIG. 3A), the differential pressure P, the position X of the positioning target 10, and the position deviation. It is a graph which shows the time change of e. In the fourth graph, the position X of the positioning target 10 is shown by a solid line, and the position deviation e is shown by a broken line. The control cycle Tc is a cycle in which an operation pulse is given.

In the comparative example shown in FIG. 4, a position command for changing the command position Xref by the step amount ΔXstp was given at time t0, and the position correction operation was performed by 10 operation pulses. The control device 50 of the hydraulic drive position control system according to the comparative example does not include the friction force compensating unit 52 (FIG. 3A). Therefore, only the first pulse width Tw1 corresponding to the position deviation e is input to the pulse generation unit 53. The electric motor 22 is driven for a time corresponding to the first pulse width Tw1 for each control cycle Tc. When the command position Xref changes by the step amount ΔXstp at time t0, the position deviation e changes from zero to ΔXstp.

The deviation / pulse width conversion unit 51 (FIG. 3A) outputs the first pulse width Tw1 corresponding to the position deviation e, and the hydraulic pump 21 is driven for a time corresponding to the first pulse width Tw1. The hydraulic oil corresponding to the first pulse width Tw1 flows into the first oil chamber 13A of the hydraulic cylinder 11, and the pressure in the first oil chamber 13A rises. The hydraulic oil in the second oil chamber 13B is sucked out by the hydraulic pump 21 and the pressure drops. In this way, a pressure difference (differential pressure P) between the first oil chamber 13A and the second oil chamber 13B is generated. When the hydraulic pump 21 stops, the supply of hydraulic oil from the hydraulic pump 21 to the hydraulic cylinder 11 stops, and the pressure rise in the first oil chamber 13A also stops. Further, due to the leakage from the hydraulic pump 21, a part of the hydraulic oil in the first oil chamber 13A, the second oil chamber 13B and the piping leaks, and the pressure gradually drops. As a result, the differential pressure P also gradually drops. However, since the amount of step-down due to leakage is smaller than the amount of step-up of pulse drive, the differential pressure P rises in a sawtooth shape each time the pulse drive is repeated.

Until the differential pressure P of the hydraulic cylinder 11 reaches the friction equivalent pressure Pfrc, the positioning target 10 does not move, and the position deviation e does not change with the step amount ΔXstp. Therefore, the first pulse width Tw1 is equal to the initial pulse width, and the amount of hydraulic oil supplied from the hydraulic pump 21 to the hydraulic cylinder 11 is also constant. As the number of operation pulses increases, the total amount of hydraulic oil supplied to the hydraulic cylinder 11 increases, and the differential pressure P gradually increases.

When the peak value of the differential pressure P exceeds the friction equivalent pressure Pfrc, the positioning target 10 starts to move. In the example shown in FIG. 4, the peak value of the differential pressure P exceeds the friction equivalent pressure Pfrc at time t1, and the positioning target 10 starts moving.

After the positioning target 10 starts moving, the position X approaches the command position Xref, so that the position deviation e gradually decreases. The first pulse width Tw1 becomes shorter as the position deviation e decreases. In the example shown in FIG. 4, the position X coincides with the command position Xref at time t2, and the position deviation e becomes zero.

In the comparative example shown in FIG. 4, the positioning target 10 starts moving at time t1. The time from time t0 to t1 is the wasted time Td in which the positioning target 10 does not move at all.

FIG. 5 shows a position command when the positioning target 10 (FIG. 1) is moved to the command position Xref using the hydraulic drive position control system according to the embodiment, an operation pulse generated by the pulse generating unit (FIG. 3A), and a differential pressure P. , And the time change of the position X and the position deviation e of the positioning target 10. In the fourth graph, the position X of the positioning target 10 is shown by a solid line, and the position deviation e is shown by a broken line.

In the embodiment shown in FIG. 5, the command position Xref changes by the step amount ΔXstp at time t0, and the response waveform obtained by performing the position correction operation by the pulse output six times is shown. When the position command of the step amount ΔXstp is given at the time t0, the position deviation e becomes equal to the step amount ΔXstp, and the second pulse width TW2 becomes equal to Tfrc from the equation (7). However, it is assumed that the differential pressure P is zero.

The third pulse width Tw3 obtained by adding the second pulse width Tw2 to the first pulse width Tw1 is input to the pulse generation unit 53 (FIG. 3A), and the amount of hydraulic oil supplied to the hydraulic cylinder 11 is a comparative example of FIG. From the case of, the amount is increased by the amount corresponding to the second pulse width TW2. As a result, the peak value of the differential pressure P generated at the first pulse output at time t0 also becomes higher than the peak value shown in the comparative example of FIG.

Since the differential pressure P is larger than the differential pressure P at the start of the control cycle at the immediately preceding time t0 at the start of the second control cycle at time t3, the second pulse width Tw2 obtained from the equation (7) is , It becomes shorter than the second pulse width Tw2 obtained in the immediately preceding control cycle. However, since Pfrc-P is a positive value, the second pulse width Tw2 has a positive value. Therefore, the amount of hydraulic oil supplied to the hydraulic cylinder 11 in the control cycle at time t3 is the amount of hydraulic oil supplied to the hydraulic cylinder 11 in the control cycle next to the control cycle at time t0 in the comparative example of FIG. is more than.

In the example shown in FIG. 5, the peak value of the differential pressure P exceeds the friction equivalent pressure Pfrc at the timing of time t4 between the second control cycle and the third cycle, and the positioning target 10 starts to move.

After the positioning target 10 starts moving, the position X approaches the command position Xref, and the position deviation e approaches zero. When the differential pressure P becomes large and the position deviation e becomes small, the second pulse width Tw2 obtained from the equation (7) suddenly becomes small and the effect of friction compensation weakens, and the first pulse width corresponding to the position deviation e is substantially obtained. The electric motor 22 is driven based on Tw1. At time t5, the position deviation e becomes zero.

As described above, in this embodiment, the third pulse width Tw3, which is the operation pulse at the start of the first control cycle after the command position Xref is changed, is larger than that in the comparative example shown in FIG. Therefore, it is possible to reduce the wasted time Td from the stationary state to the actual start of the positioning target 10.

After the differential pressure P exceeds the friction equivalent pressure Pfrc and the positioning target 10 starts moving, it is preferable that the compensation function by the friction force compensating unit 52 does not work. When the positioning target 10 starts to move, the difference between the differential pressure P and the friction equivalent pressure Pfrc in the equation (7) and the position deviation e gradually decrease. Therefore, the second pulse width Tw2 generated by the frictional force compensating unit 52 is shortened. Therefore, when the positioning target 10 starts to move, the influence of the frictional force compensating unit 52 on the position control becomes small. In this way, the frictional force compensating unit 52 exerts its effect when the influence of friction in the stationary state is high, that is, when the command position Xref changes in a stepped manner, and automatically outputs (when the positioning target 10 starts to move). It has a function of reducing the second pulse width Tw2) and eliminating the influence of frictional force compensation in the vicinity of the command position Xref.

By determining the drive time of the hydraulic pump 21 so as to generate the differential pressure P corresponding to the friction equivalent pressure Pfrc when the positioning target 10 is in a stationary state, the positioning target 10 is from stationary to the start of movement. Wasted time can be reduced. However, this method is valid when the differential pressure of the stationary hydraulic cylinder is zero, but becomes a problem when the differential pressure is not zero and is biased to positive or negative. For example, when the differential pressure P is positively biased, the difference between the friction equivalent pressure Pfrc and the differential pressure P becomes small, and the amount of hydraulic oil is too large at the pulse width of Tfrc, and the differential pressure P exceeds the friction equivalent pressure Pfrc. Will move. Since the pulse width of Tfrc is larger than the range of the first pulse width Tw1, there is a possibility that the command position Xref may be jumped over and positioning at the command position Xref may not be possible in some cases. On the contrary, when the differential pressure P is negatively biased, the amount of hydraulic oil is insufficient, the differential pressure P cannot reach the friction equivalent pressure Pfrc, and the pressure is increased after the second pulse, resulting in wasted time. The result is. As described above, it is important to output an appropriate pulse width according to the difference between the pressure state of the hydraulic cylinder 11 and the friction equivalent pressure Pfrc.

In the method of generating the differential pressure P corresponding to the friction equivalent pressure Pfrc when the positioning target 10 is in a stationary state, the friction equivalent pressure Pfrc is not considered during the movement of the positioning target 10 and is based on the position deviation e. The drive time of the hydraulic pump 21 will be determined. Therefore, it is necessary to accurately determine whether the frictional force acting on the positioning target 10 is a static frictional force or a dynamic frictional force when moving at a very small speed. However, it is difficult to accurately distinguish between these two states. If you make a mistake in distinguishing between the two, control will become unstable.

In this embodiment, the same control may be performed when the positioning target 10 is stationary and when it is moving. Therefore, it is not necessary to determine whether the frictional force acting on the positioning target 10 is a static frictional force or a dynamic frictional force when moving at a very small speed. Therefore, stable control can be performed.

Next, a modified example of the above embodiment will be described.
In the above embodiment, the piston 11P, the first rod 11A, and the second rod 11B of the hydraulic cylinder 11 (FIG. 1) are fixed and the cylinder body 11S is moved, but conversely, the cylinder body 11S is fixed. The piston 11P, the first rod 11A, and the second rod 11B may be moved together with the positioning target 10. Further, in the above embodiment, the double-rod hydraulic cylinder is used as the hydraulic cylinder 11, but other hydraulic cylinders such as a double-acting single-rod hydraulic cylinder and a single-acting single-rod hydraulic cylinder may be used.

In the above embodiment, the hydraulic cylinder 11 that moves the positioning target 10 along a straight line is used, but other hydraulic actuators may be used instead of the hydraulic cylinder 11. For example, a hydraulic motor may be used as the hydraulic actuator. In this case, the control device 50 of the above embodiment may be applied to the hydraulic drive position control system that controls the position of the hydraulic motor in the rotational direction.

In the above embodiment, the control device 50 gives the angular velocity ω and the drive time Δt of the equation (4) to the motor driver 25 by the operation pulse. As a modification thereof, the control device 50 may give the motor driver 25 information for designating the rotation angle Δθ of the equation (4). At this time, the motor driver 25 rotates the electric motor 22 by a given rotation angle Δθ.

It goes without saying that the above-described examples and modifications are examples, and partial replacement or combination of the configurations shown in different examples and modifications is possible. Similar effects and effects due to the same configuration of the plurality of examples and modifications will not be mentioned sequentially for each of the examples and modifications. Furthermore, the present invention is not limited to the above-mentioned examples and modifications. For example, it will be obvious to those skilled in the art that various changes, improvements, combinations, etc. are possible.

10 Positioning target 10A Guide sliding surface (convex)
11 Hydraulic cylinder 11A 1st rod 11B 2nd rod 11P Piston 11S Cylinder body 12A 1st oil passage 12B 2nd oil passage 13A 1st oil chamber 13B 2nd oil chamber 14A 1st port 14B 2nd port 15A, 15B Pressure sensor 16 Position sensor 18 Support member 18A Guide sliding surface (concave)
20 Hydraulic oil supply system 21 Hydraulic pump 22 Electric motor 23A, 23B Relief valve 24 Shuttle valve 25 Motor driver 26 Charge pump 27 Electric motor 28 Relief valve 30 Hydraulic oil tank 50 Control device 51 Deviation / pulse width conversion unit 52 Friction force compensation unit 53 Pulse generator

Claims (7)

A control device that drives a hydraulic actuator by controlling a hydraulic oil supply system that supplies hydraulic oil to a hydraulic actuator that generates thrust to move a positioning object.
The hydraulic oil is supplied from the hydraulic oil supply system to the hydraulic actuator at each control cycle.
A hydraulic drive position control device that increases the supply amount of hydraulic oil in the first control cycle to be larger than the supply amount of hydraulic oil in the second and subsequent control cycles during the period from the stationary state to the start of movement of the positioning target. The hydraulic drive position control device according to claim 1, wherein the supply amount of the hydraulic oil is changed by adjusting the time for supplying the hydraulic oil from the hydraulic oil supply system to the hydraulic actuator for each control cycle. The hydraulic drive position control device according to claim 1 or 2, wherein a plurality of control cycles are executed until the positioning target reaches a command position based on a measured value of the position of the positioning target. The first operation amount is generated according to the deviation between the command position of the positioning target and the measured value of the positioning target position.
A second manipulated variable is generated according to the difference between the measured value of the physical quantity according to the thrust generated by the hydraulic actuator and the set value of the physical quantity that defines the preset thrust.
The hydraulic drive position control device according to claim 3, wherein the hydraulic oil supply system is controlled based on the third operation amount obtained by adding the second operation amount to the first operation amount. The hydraulic drive position control device according to claim 4, wherein when the command position of the positioning target changes, the second operation amount is generated based on the ratio of the deviation to the latest change amount of the command position. A control device that drives a hydraulic actuator by controlling a hydraulic pump that supplies hydraulic oil to a hydraulic actuator that generates thrust to move a positioning target.
Hydraulic oil is supplied from the hydraulic pump to the hydraulic actuator at each control cycle.
The period from the stationary state to the start of movement of the positioning target is based on the load pressure acting on the pressure receiving surface of the hydraulic actuator in the first pulse width corresponding to the position deviation between the command position of the positioning target and the current position. A hydraulic drive position control device that drives the hydraulic pump for a time corresponding to the third pulse width to which the second pulse width is added and supplies hydraulic oil to the hydraulic actuator. A claim for determining the second pulse width based on the difference between the friction equivalent pressure in which the maximum static friction force acting on the positioning target is converted into a load pressure and the load pressure acting on the pressure receiving surface of the hydraulic actuator. Item 6. The hydraulic drive position control device according to Item 6.

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