JP2011038558A - Electric fluid pressure actuator device - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide an electric fluid pressure actuator device in which iron loss of a motor is suppressed without causing complexity in constitution while being capable of fulfilling an original function of a pressure compensation mechanism without limitation so as to reduce power consumption of the motor for holding the position of an output portion. <P>SOLUTION: A closed flow passage is formed of a cylinder 12 having a piston 15, and a swash plate variable displacement pump 13. The pressure compensation mechanism 20 is provided to reduce the displacement of the pump 13 according to the differential pressure when a differential pressure between pressures P1 and P2 is not less than a set value Pc. A control device 19 controls a rotational speed and a rotational direction of a servo motor 14 for driving the pump 13 in accordance with an input position instruction signal and a position detection signal of the piston 15 by a position sensor 16. At that time, when the piston 15 is positioned in the vicinity of the target position (reaching 95% or more of the whole movement distance to a target position), the rotational speed of the servo motor 14 is restricted to a predetermined value or less. <P>COPYRIGHT: (C)2011,JPO&INPIT

Description

  The present invention relates to an electric fluid pressure actuator device having a structure in which a variable displacement fluid pressure pump and a servo motor are integrated with a fluid pressure actuator.

  For example, as an actuator used in, for example, transportation equipment and a press device, in recent years, an electric fluid abbreviated as EHA (Electro Hydrostatic Actuator) including a fluid pressure cylinder, a fluid pressure pump, and a servo motor instead of the conventional hydraulic type. Pressure actuator devices have been provided. This EHA has advantages such that a long hydraulic pipe is not required, the whole is small and light, and maintenance is good.

  FIG. 7A shows the basic configuration (first conventional example) of this type of EHA. That is, the hydraulic cylinder 1 is divided into two chambers, a left chamber 1a and a right chamber 1b, by a piston 2, and two inlets and outlets of a hydraulic pump 3 capable of discharging and sucking hydraulic oil from both directions are connected to the left chamber 1a and the left chamber 1a. Each is connected to the right chamber 1b. The motor 4 that drives the hydraulic pump 3 is controlled by the controller 5 such as the rotation speed N, the rotation direction, and the like. The controller 5 receives a position command signal Ecom and a feedback signal Exp from a position sensor 6 that detects the position of the piston 2 (piston rod 2a).

  Thus, the controller 5 outputs a command corresponding to the difference between the position command signal Ecom and the feedback signal Exp, the motor 4 rotates at the rotation speed Nm according to the command, and the pump 3 is qth ( A discharge capacity per revolution) × Nm of flow rate Q is generated. When the flow rate Q is generated in the direction of the arrow in FIG. 7A, the hydraulic oil flows into the left chamber 1a of the cylinder 1, and the piston 2 moves in the direction of the arrow Xp. At this time, the hydraulic oil corresponding to the volume reduction of the right chamber 1b is sucked into the pump 3. In a state where the position command signal Ecom and the feedback signal Exp match, the command to the motor 4 becomes 0, and a position control loop is formed.

  Further, the behavior of the position command Ecom, the displacement Xp of the piston 2 and the rotational speed N of the motor 4 with the passage of time at this time is as shown in FIG. 7A illustrates the case where the position sensor 6 is arranged coaxially with the piston 2. However, as shown in FIG. 7B, the position sensor 6 ′ is arranged in parallel with the piston 2. There may be cases.

  By the way, in the configuration of the EHA, when an external force (external load) F in the direction of the white arrow acts on the piston 2, the pressure P1 in the left chamber 1a of the cylinder 1 and the pressure P2 in the right chamber 1b If the pressure difference ΔP that satisfies P1> P2 does not occur in the meantime, the position Xp of the piston 2 cannot be held at the command position. The differential pressure ΔP at this time needs to be a value satisfying Acy × (P1−P2) = F, where Acy is the pressure receiving area of the piston 2. In an ideal state where there is no leakage in the hydraulic system, the differential pressure ΔP is maintained by the pump 3 continuously pushing out the hydraulic oil with a torque of qth × (P1−P2) = Tp. In order to continue to generate this torque Tp, it is necessary to continue to pass a current through the coil of the motor 4 while keeping the discharge flow rate of the pump 3 at zero. For this reason, there is a problem that power consumption for maintaining the position of the piston 2 becomes relatively large.

  In order to solve the above problem, for example, Patent Document 1 proposes to provide a pressure compensation mechanism in the EHA (second conventional example). As shown in FIG. 9, this pressure compensation mechanism is a well-known swash plate type variable capacity capable of changing the discharge amount by changing the inclination angle of the swash plate by a swash plate control mechanism instead of the pump 3. A type pump 8 is employed, and a shuttle valve 9 is provided to guide the higher pressure of the pressure P1 in the left chamber 1a and the pressure P2 in the right chamber 1b of the cylinder 1 to the swash plate control mechanism.

  At this time, as shown in FIG. 10, when the differential pressure ΔP (difference between the pressure P1 and the pressure P2) becomes equal to or greater than Pc, the pressure compensation mechanism is activated and the discharge capacity of the pump 8 is reduced. As a result, the torque that must be generated by the motor 4, and hence the current, is reduced, so that power consumption can be suppressed. Actually, since the hydraulic oil leaks in the pump 8, it is necessary to provide a tank (accumulator) and a second shuttle valve to supply hydraulic oil to the hydraulic system.

Japanese Patent No. 4022032

  However, in the EHA provided with the pressure compensation mechanism as described above, the following problems are expected to occur. That is, in the EHA provided with the pressure compensation mechanism shown in FIG. 9, the behavior of the position command Ecom, the displacement Xp of the piston 2, the rotational speed N of the motor 4 and the swash plate angle α with the passage of time is shown in FIG. As shown. Now, when the external force F acts on the piston 2 and the differential pressure ΔP exceeds Pc (time t1), the swash plate angle α is reduced and the load torque of the motor 4 is reduced. At this time, the TN characteristic of the motor 4 is as shown in FIG. 12, and the rotational speed N of the motor 4 increases as the load torque T decreases.

  In this case, when a large external force F is applied such that the differential pressure ΔP is close to P0, the discharge amount of the pump 8 approaches zero. Therefore, in order to maintain the position of the piston 2, the leakage of the pump 8 is compensated. The rotational speed N of the motor 4 increases to the maximum rotational speed. As a result, since the capacity of the pump 8 is reduced, the required torque, that is, the required current of the motor 4 is reduced, but the iron loss of the motor 4 is increased, and as a result, the insulation of the winding is damaged by the heat generation of the motor 4. Problems such as receiving.

  In order to solve the problem of the iron loss of the motor 4, it is conceivable to limit the swash plate angle α (the angle is not changed to the minimum swash plate angle α). According to this, since a certain amount of load torque is applied to the motor 4 and the necessary rotational speed of the motor 4 for compensating for the leakage of the pump 8 can be reduced, the rotational speed N of the motor 4 is reduced. It will not increase to the maximum value, and iron loss can be suppressed. However, in this method, since the reduction of the load torque of the motor 4 is limited, the power consumption of the motor 4 increases accordingly. At the same time, since the original pressure compensation function does not work, it is necessary to add a relief valve to control the upper limit pressure of the system (to release excess hydraulic fluid), which complicates the configuration and consumes unnecessary energy. I will invite you.

  The present invention has been made in view of the above circumstances, and an object of the present invention is to use without limitation the original function of the pressure compensation mechanism for minimizing the power consumption of the motor for maintaining the position of the output unit. However, it is an object of the present invention to provide an electric fluid pressure actuator device that can suppress the iron loss of the motor without complicating the configuration.

  In order to achieve the above object, an electric fluid pressure actuator device of the present invention forms a closed flow path between a fluid pressure actuator having an output part that performs reciprocating linear motion or rotational motion, and the fluid pressure actuator. A variable displacement fluid pressure pump, a servo motor for driving the fluid pressure pump, a position sensor for detecting the position of the output portion of the fluid pressure actuator, and the fluid pressure actuator based on the detection of the position sensor. And a control device that controls the servo motor so as to move the output unit to a target position and maintain the position, the differential pressure between the discharge side and the suction side of the fluid pressure pump is a set value. A pressure compensation mechanism is provided for reducing the capacity of the fluid pressure pump according to the magnitude of the differential pressure when the pressure exceeds Pc, and the fluid pressure When the output unit of the actuator is positioned near the target position is characterized at which a limit means for limiting the rotational speed of the servo motor to a predetermined value or less.

  According to the said structure, the output part of a fluid pressure actuator can be moved to arbitrary positions by controlling the servomotor which drives a fluid pressure pump with a control apparatus, The position can be hold | maintained. When the output portion of the fluid pressure actuator receives an external force and the pressure difference between the discharge side and the suction side of the fluid pressure pump becomes equal to or larger than the set value Pc, the pressure compensation mechanism causes the variable displacement type fluid pressure. Since the discharge amount of the pump is reduced, the current of the servo motor is reduced and the power consumption can be suppressed.

  Here, when the discharge capacity of the fluid pressure pump is reduced by the pressure compensation mechanism, the load torque of the servo motor is reduced. However, in the present invention, when the output portion of the fluid pressure actuator is located in the vicinity of the target position. The limiting means limits the rotational speed of the servo motor to a predetermined value or less. In this case, in a general control system, as shown in FIG. 8, when the output unit (piston) reaches the target position, the rotation speed of the servo motor becomes zero. That is, since the rotation speed of the servo motor is reduced near the target position, there is no real harm even if the rotation speed of the servo motor is limited to a small value near the target position.

  And when the rotation speed of the servo motor becomes low and the discharge amount of the fluid pressure pump is smaller than the flow rate required by the system, the system pressure decreases. The discharge amount is increased and the supply flow rate of the fluid is increased, so that the necessary pressure is maintained. This pressure compensation mechanism does not control the position of the output part, but if the output part is insufficient with respect to the external force and the position of the output part differs from the target position (out of the range near the target position). By the control of the control device, the servo motor rotates and the position of the output unit is maintained. If a position command for a new target position is input to the control device, the current position of the output unit is no longer in the vicinity of the target position, so that the restriction on the number of rotations of the servo motor can be removed and the normal control can be restored.

  Therefore, while the original function of the pressure compensation mechanism can be used without limitation, it is possible to prevent the rotation speed of the servo motor from increasing to the maximum value as the load torque decreases. Can solve the problem of iron loss. In addition, the configuration such as the addition of a relief valve is not complicated.

  The “fluid pressure actuator” in the present invention includes a fluid pressure (hydraulic pressure, air pressure) cylinder and a fluid pressure (hydraulic pressure, air pressure) motor, and further includes a swing type actuator. In the case of a fluid pressure cylinder, a piston (piston rod) that reciprocates linearly serves as an output unit, and in the case of a fluid pressure motor, an output shaft that rotates and serves as an output unit. Further, in the present invention, the “near target position” of the output unit refers to the vicinity when, for example, 95% or more of the entire moving distance is reached from the position of the output unit at the time of input of the position command to the target position. Can be. The numerical value can be arbitrarily set and adjusted according to the control target and specifications.

  The limit number of rotations when limiting the number of rotations of the servo motor to a predetermined value or less is, for example, within a range of 1% to 50% with respect to the maximum number of rotations of the servo motor. The number of rotations can be determined. In determining the rotational speed limit at this time, it is possible to take into consideration the required conditions such as the amount of operating fluid leakage and operating characteristics of the system, and the capacity of the fluid pressure pump.

  More specifically, when a swash plate type variable displacement pump is used as the fluid pressure pump, the motor speed at the motor speed at the speed limit is that the speed limit is the maximum output and the position of the output section is maintained constant. And the power consumption for the power consumed by the pump at the stable swash plate angle at that time can be determined so as to be as small as possible. If the rotational speed limit is too low, the swash plate angle is stable at a large position, and power consumption for power increases. On the other hand, if the rotational speed limit is too high, the swash plate angle is stable at a small position and the power consumed by the power is reduced, but the power consumption such as iron loss of the motor is increased.

  According to the electric fluid pressure actuator device of the present invention, when the differential pressure between the discharge side and the suction side of the fluid pressure pump becomes equal to or larger than the set value Pc, the fluid pressure pump is selected according to the magnitude of the differential pressure. A pressure compensation mechanism is provided to reduce the capacity of the fluid pressure actuator, and when the output part of the fluid pressure actuator is located in the vicinity of the target position, a limiting means for limiting the rotation speed of the servo motor to a predetermined value or less is provided. While maintaining the original function of the pressure compensation mechanism that keeps the motor power consumption to maintain the position of the motor in an unrestricted manner, it suppresses motor iron loss without complicating the configuration. There is an excellent effect of being able to.

The circuit diagram which shows the 1st Example of this invention and shows the system configuration | structure of an electric fluid pressure actuator apparatus The flowchart which shows the control procedure with respect to the servomotor which a control apparatus performs Diagram showing the behavior of piston displacement, motor rotation speed, swash plate angle, etc. over time The figure which shows the relationship between the differential pressure by the pressure compensation mechanism and the pump discharge amount FIG. 1 equivalent view showing a second embodiment of the present invention. The figure which shows the fluid pressure cylinder part which concerns on 3rd Example of this invention. The circuit diagram which shows the 1st prior art example and shows the basic composition of an electric fluid pressure actuator device Diagram showing the behavior of piston displacement and motor rotation speed over time The circuit diagram which shows a 2nd prior art example and shows the structure of an electrohydraulic actuator device 4 equivalent diagram Diagram showing the behavior of piston displacement, motor rotation speed, swash plate angle, etc. over time Diagram showing TN characteristics of motor

  Hereinafter, a first embodiment in which the present invention is applied to an electric fluid pressure actuator apparatus having a hydraulic cylinder will be described with reference to FIGS. FIG. 1 shows the system configuration of an electric fluid pressure (hydraulic) actuator 11 (hereinafter abbreviated as EHA 11) according to this embodiment. First, the overall configuration of the EHA 11 will be described.

  The EHA 11 includes a fluid pressure (hydraulic) cylinder 12 as a fluid pressure actuator, a variable displacement fluid pressure (hydraulic) pump 13 that forms a closed flow path with the cylinder 12, and a servo motor that drives the pump 13. 14 is provided. The cylinder 12 has a piston 15 as an output portion that can reciprocate linearly in the left-right direction in the figure, and the inside of the cylinder 15 is a first chamber (left chamber) 12a and a second chamber (right chamber) 12b. It is divided into. The piston 15 has a piston rod 15a, and a load is driven by moving the piston rod 15a in the left-right direction in the figure. The position of the piston 15 (piston rod 15a) is detected by a position sensor 16.

  The pump 13 is, for example, a well-known swash plate type variable displacement piston pump, and has a first inlet / outlet port 13a and a second inlet / outlet port 13b, and is capable of discharging and sucking hydraulic oil from both directions. . At this time, although not shown in the figure, the swash plate angle α is the maximum angle and the discharge capacity qth per unit time is relatively large in the normal state. Is changed so that the discharge capacity qth decreases accordingly.

  The first inlet / outlet 13 a of the pump 13 is connected to the first chamber 12 a of the cylinder 12 via the first pipe 17, and the second inlet / outlet 13 b is connected to the second of the cylinder 12 via the second pipe 18. It is connected to the chamber 12b. The servo motor 14 for driving the pump 13 is controlled in its rotational speed N, rotational direction, and the like by a control device 19 including a CPU. The maximum rotation speed of the servo motor 14 is set to 6000 rpm, for example.

  The control device 19 is supplied with a position command signal Ecom and a feedback signal (position detection signal) Exp from the position sensor 16. The control device 19 outputs a command corresponding to the difference between the position command signal Ecom and the feedback signal Exp to the servo motor 14. The servo motor 14 rotates at a rotation speed Nm according to the command, and the pump 13 generates a flow rate Q of qth (discharge capacity per rotation) × Nm.

  For example, when the flow rate Q is generated in the direction of the arrow in FIG. 1, the hydraulic oil flows into the first chamber 12a of the cylinder 12 (outflow from the second chamber 12b), and the piston 15 moves in the direction of the arrow Xp. At this time, hydraulic oil corresponding to the volume reduction of the second chamber 12 b of the cylinder 12 is sucked into the pump 13. When the piston 15 moves to the target position and the position command signal Ecom and the feedback signal Exp coincide with each other, the command to the motor 14 becomes 0, and a position control loop is formed.

  In addition, when an external force (external load) F in the direction of the white arrow acts on the piston 15, P1> between the pressure P1 in the first chamber 12a of the cylinder 12 and the pressure P2 in the second chamber 12b. If the differential pressure ΔP of P2 does not occur, the position Xp of the piston 15 cannot be held at the target position. The differential pressure ΔP at this time is a value satisfying Acy × (P1−P2) = F, where Acy is the pressure receiving area of the piston 15. This differential pressure ΔP is maintained by the pump 13 continuing to push out the hydraulic oil with a torque of qth × (P1−P2) = Tp. In order to continue to generate this torque Tp, it is necessary to keep the current flowing through the coil of the motor 14 while keeping the discharge flow rate of the pump 13 at zero (actually, the discharge flow rate is generated by the amount of system leakage). Power consumption will increase.

  Therefore, in the EHA 11 of this embodiment, when the pressure difference between the discharge side and the suction side of the pump 13 (the differential pressure ΔP between the pressures P1 and P2) is equal to or greater than the set value Pc (see FIG. 4). A pressure compensation mechanism 20 for reducing the capacity of the pump 13 in accordance with the magnitude of the differential pressure ΔP is provided. In this case, the pressure compensation mechanism 20 controls the pressure in the piston chamber 21a of the swash plate control piston 21 for changing the swash plate angle α of the pump 13 and the swash plate control piston 21, as shown in FIG. The first shuttle for selectively taking the higher pressure of the swash plate control valve 22, the first pipe 17 (pressure P1) and the second pipe 18 (pressure P2) into the swash plate control valve 22. A valve 23 is provided.

  Further, as shown in FIG. 1, the EHA 11 includes a pressurized tank (accumulator) 24 and an accumulator 24 in order to replenish hydraulic oil lost due to leakage in the closed hydraulic system between the pump 13 and the cylinder 12. A second shuttle valve 25 is provided. The second shuttle valve 25 is provided in parallel with the first shuttle valve 23 between the first pipe 17 and the second pipe 18, and selects the lower one of the pressure P1 and the pressure P2, The hydraulic oil is supplied from the pressurized tank 24 to the hydraulic system. The pressurized tank 24 is also connected to the casing of the pump 13.

  In the pressure compensation mechanism 20, when the differential pressure ΔP between the pressure P1 and the pressure P2 is lower than the set differential pressure Pc, the valve body 22a of the swash plate control valve 22 is closed, and as shown in FIG. The piston chamber 21 a of the plate control piston 21 is connected to the pressurized tank 24. In this state, the swash plate angle α of the pump 13 is maintained at the maximum by the spring of the swash plate control piston 21, and the discharge amount Q of the pump 13 is also maximized. The relationship between the differential pressure ΔP and the discharge amount Q of the pump 13 at this time is as shown in FIG. 4, and the discharge amount Q becomes the maximum value (a constant value) until the differential pressure ΔP reaches the set value Pc. . The pressure compensation mechanism 20 is set so that the discharge amount Q gradually decreases toward P0 when the differential pressure ΔP becomes equal to or greater than the set value Pc.

  In FIG. 1, the differential pressure applied to the swash plate control valve 22 is the difference between the higher pressure of the pressures P1 and P2 and the pressure of the oil passage of the accumulator 24 system, and the differential pressure between P1 and P2 It is not. Actually, the lower pressure of the pressures P1 and P2 by the shuttle valve 25 and the pressure in the oil passage of the accumulator 24 system coincide with each other. Therefore, the differential pressure applied to the swash plate control valve 22 is the differential pressure between P1 and P2. Matches.

  When the differential pressure ΔP reaches the set differential pressure Pc, the valve element 22a of the swash plate control valve 22 is displaced to the right in the figure against the spring force, and the differential pressure is transferred to the piston chamber 21a of the swash plate control piston 21. Introduced, the swash plate control piston 21 is pushed down to reduce the swash plate angle α. Thereby, the discharge capacity of the pump 13 is reduced. When the pump capacity becomes too small and the discharge amount Q falls below the system required flow rate, the valve body 22a of the swash plate control valve 22 is returned by the spring force, and the swash plate angle α of the pump 13 increases. Thus, when the differential pressure ΔP is between Pc and P0, as shown in FIG. 4, the system flow rate including leakage and the discharge amount of the pump 13 are balanced. It should be noted that the slope of the decrease in flow rate between Pc and P0 in FIG. 4 can be set sharply or gently as required by the design of the swash plate control system.

  In the present embodiment, as will be described in the following description of the operation (flowchart description), the control device 19 controls the servomotor 14 with the software configuration based on the detection of the position sensor 16 based on the detection of the position sensor 16. When the twelve pistons 15 are located in the vicinity of the target position, the rotational speed N of the servo motor 14 is limited to a predetermined value or less. Therefore, the control device 19 functions as a limiting unit.

  At this time, specifically, in the vicinity of the target position of the piston 15, for example, in a state where it has reached 95% or more of the entire moving distance from the position of the piston 15 when the position command is input to the target position, It is determined that the vicinity of the target position has been reached.

  In addition, when limiting the rotational speed N of the servo motor 14 to a predetermined value or less, the rotational speed limit (predetermined value) is the maximum rotational speed at which the piston 15 is maintained at a constant rotational speed. The sum of the motor loss and the power consumption for the power consumed by the pump 13 at the stable swash plate angle at that time can be determined to be as small as possible. More specifically, it is desirable to determine a predetermined number of rotations within a range of, for example, 1% or more and 50% or less with respect to the maximum number of rotations of the servo motor 14. In the present embodiment, for example, the maximum rotational speed of the servo motor 14 is 600 rpm, whereas the limiting rotational speed is 1000 rpm.

  Next, the operation of the above configuration will be described with reference to FIGS. In the EHA 11 configured as described above, the control device 19 controls the servo motor 14 that drives the pump 13 based on the position command signal Ecom and the feedback signal Exp, and moves the piston 15 of the cylinder 12 to an arbitrary target position. The position can be held. At this time, as described below, when the piston 15 of the cylinder 12 is located in the vicinity of the target position (95% or more of the moving distance to the target position), the control device 19 sets the rotation speed of the servo motor 14 to a predetermined value. Control is performed so as to limit to a value (1000 rpm) or less.

  Further, in the EHA 11, since the pressure compensation mechanism 20 is provided, the piston 15 of the cylinder 12 receives the external force F and the pressure difference between the discharge side and the suction side of the pump 13 (the difference between the pressures P1 and P2). When the pressure ΔP) is equal to or higher than the set value Pc, the swash plate angle α is reduced and the discharge amount of the pump 13 is reduced. Therefore, the current consumption of the servo motor 14 can be reduced to reduce power consumption.

  The flowchart of FIG. 2 shows a control procedure of the servo motor 14 that is executed by the control device 19 when the position command signal Ecom is input. That is, first, in step S1, it is determined whether or not the position deviation Eer obtained from the position command signal Ecom and the feedback signal Exp is not 0 (greater than 0). The position deviation Eer at this time is obtained by multiplying the value of (Ecom-Exp) by a predetermined adjustment amount.

  If the position deviation Eer is larger than 0 (Yes in step S1), in the next step S2, a command (specifically, a rotational direction and a rotational speed of the servo motor 14 proportional to the size of the position deviation Eer is specified. Motor applied voltage) is output to the servo motor 14. In the next step S3, it is determined whether the feedback signal (position detection signal) Exp from the position sensor 16 is 95% or more of the position command signal Ecom, that is, whether the piston 15 is positioned in the vicinity of the target position.

  When it is determined that the piston 15 is located in the vicinity of the target position (Yes in step S3), the maximum rotational speed of the servo motor 14 is limited to 1000 rpm in step S4, and the process returns to step S1. If it is determined in step S3 that the piston 15 is not located in the vicinity of the target position (No in step S3), the rotational speed of the servo motor 14 is not limited, and the process returns to step S1 as it is. If it is determined in step S1 that the position deviation Eer is 0 (No in step S1), the motor command is set to 0 (stop) in step S5, and the process returns to step S1.

  By the control as described above, in this embodiment, the behavior of the piston displacement Xp, the motor rotation speed N, the swash plate angle α, and the like with the passage of time is as shown in FIG. If the differential pressure ΔP exceeds the set value Pc at time t1, the swash plate angle α is decreased by the action of the pressure compensation mechanism 20, the discharge capacity of the pump 13 is reduced, and the load torque of the servo motor 14 is decreased. T decreases. At this time, as shown in FIG. 12, in the servo motor 14, the rotational speed N increases as the load torque T decreases. In the EHA of the second conventional example described above, as shown in FIG. 11, the pressure compensation mechanism minimizes the swash plate angle α and increases the motor rotation speed N to the maximum rotation speed. There were problems such as iron loss.

  However, in this embodiment, when the piston 15 is positioned in the vicinity of the target position, that is, when the displacement Xp of the piston 15 reaches 95% of the target position (time t2), the rotational speed of the servo motor 14 is limited (1000 rpm or less). ) Can be applied. For this reason, unlike the behavior shown in FIG. 11, the swash plate angle α does not decrease to the minimum, but returns to the angle corresponding to the limit rotational speed of the servo motor 14 and the leakage of hydraulic fluid of the system. There is no need to limit the function of the pressure compensation mechanism 20 (for example, to provide a lower limit of the swash plate angle).

  In this case, in a general control system, when the piston 15 reaches the target position, the rotation speed of the servo motor 14 becomes zero. That is, since the rotation speed of the servo motor 14 is small near the target position, there is no real harm even if the rotation speed of the servo motor 14 is limited to a small value near the target position. Although not shown, when the rotation speed of the servo motor 14 is low and the discharge amount of the pump 13 is smaller than the flow rate required by the system, the system pressure decreases. As a result, the swash plate angle α is increased, the discharge amount of the pump 13 is increased, the supply flow rate of hydraulic oil is increased, and the necessary pressure is maintained.

  Although the pressure compensation mechanism 20 does not control the position of the piston 15, the force of the piston 15 is insufficient with respect to the external force F, and the position of the piston 15 differs from the target position (out of the range near the target position). Then, under the control of the control device 19, the rotational speed of the servo motor 14 is increased and the position of the piston is maintained. If the position command Ecom of a new target position is input to the control device 19, the current position of the piston 15 is no longer near the target position, so that the limitation on the rotation speed of the servo motor 14 is removed and the normal control can be restored. is there.

  Thus, according to the present embodiment, without limiting the change angle of the swash plate angle α, the rotation speed of the servo motor 14 is increased to the maximum value as the load torque is reduced by limiting the rotation speed of the servo motor 14. It is possible to prevent a rise to the level. Therefore, according to the EHA 11 of this embodiment, the original function of the pressure compensation mechanism 20 can be used without limitation, but the problem of iron loss of the servo motor 14 that has been a conventional concern can be solved. . In addition, it is not necessary to add a relief valve for controlling the upper limit pressure of the system (to release excess hydraulic oil), and the configuration is not complicated and wasteful energy is not consumed.

  FIG. 5 shows a system configuration of the electric hydraulic actuator device 31 according to the second embodiment of the present invention. The second embodiment differs from the first embodiment in that a fluid pressure (hydraulic) motor 32 is employed as the fluid pressure actuator instead of the fluid pressure cylinder 12. The fluid pressure motor 32 has two inlet / outlet ports 32a and 32b that flow in both directions. One of the inlet / outlet ports 32a is connected to the first inlet / outlet port 13a of the pump 13 via the first pipe 17, and the other inlet / outlet port 32b is connected to the first inlet / outlet port 32b. It is connected to the second inlet / outlet 13 b of the pump 13 through the two pipes 18.

  The fluid pressure motor 32 has an output shaft 33 as an output unit that performs rotational movement, and a signal from a position sensor 34 (rotation sensor) that detects the position of the output shaft 33 is input to the control device 19. The Other configurations, that is, the point provided with the pressure compensation mechanism 20 and the point that the control device 19 functions as the limiting means are the same as those in the first embodiment. Even in the second embodiment, the original function of the pressure compensation mechanism 20 for minimizing the power consumption of the servo motor 14 for maintaining the position of the output shaft 33 can be used without limitation. The same effects as those of the first embodiment can be obtained, such as suppressing iron loss of the servo motor 14 without increasing the complexity.

  Further, as in the third embodiment of the present invention shown in FIG. 6, a so-called single rod type fluid pressure (hydraulic) cylinder 41 may be employed as the fluid pressure actuator. The cylinder 41 is divided into a first chamber 41a and a second chamber 41b by a piston 42 as an output part that performs a reciprocating linear motion, but the piston rod 42a exists only on the second chamber 41b side. The first chamber 41a side does not exist. The position of the piston 42 (piston rod 42a) is detected by the position sensor 43.

In addition, the present invention is not limited to the above-described embodiments, and various expansions and modifications as described below, for example, are possible.
In other words, the limiting rotational speed when the rotational speed of the servo motor 14 is limited to a predetermined value or less can be freely taken into consideration, for example, required conditions such as the amount of system hydraulic fluid leakage and the capacity of the fluid pressure pump 13. Can be determined. At this time, when a swash plate type variable displacement pump is used, the motor speed at the speed limit and the stable swash plate angle at that time with the maximum speed output and the piston position kept constant. Thus, the total power consumption for the power consumed by the pump can be determined to be as small as possible. If the rotational speed limit is too low, the swash plate angle is stable at a large position, and power consumption for power increases. On the other hand, if the rotational speed limit is too high, the swash plate angle is stable at a small position and the power consumed by the power is reduced, but the power consumption such as iron loss of the motor is increased.

  In the first embodiment, when the displacement of the piston 15 reaches 95% or more with respect to the entire moving distance to the target position, it is determined that the piston 15 is positioned in the vicinity of the target position. Can be arbitrarily set and adjusted in accordance with, for example, the control target and specifications. In addition, various modifications can be made to the configuration of the variable displacement fluid pressure pump and the specific configuration of the pressure compensation mechanism. It is possible.

  In the drawings, 11 is an EHA (electric fluid pressure actuator device), 12 and 41 are fluid pressure cylinders (fluid pressure actuators), 13 is a fluid pressure pump, 14 is a servo motor, 15 and 42 are pistons (output units), 16, 34 and 43 are position sensors, 19 is a control device (limitation means), 20 is a pressure compensation mechanism, 31 is an electric fluid pressure actuator device, 32 is a fluid pressure motor (fluid pressure actuator), and 33 is an output shaft (output unit). Show.

Claims (1)

A fluid pressure actuator having an output part for reciprocating linear motion or rotational motion;
A variable displacement fluid pressure pump that forms a closed channel with the fluid pressure actuator;
A servo motor that drives the fluid pressure pump;
A position sensor for detecting the position of the output part of the fluid pressure actuator;
In the electric fluid pressure actuator device comprising: a control device for controlling the servo motor so as to move the output portion of the fluid pressure actuator to a target position based on the detection of the position sensor and hold the position;
When the pressure difference between the discharge side and the suction side of the fluid pressure pump exceeds a set value Pc, a pressure compensation mechanism is provided that reduces the capacity of the fluid pressure pump according to the magnitude of the pressure difference. With
An electric fluid pressure actuator device comprising a limiting means for restricting the rotation speed of the servo motor to a predetermined value or less when the output portion of the fluid pressure actuator is located in the vicinity of a target position.

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