JP6924043B2 - Vibration damping device for railway vehicles - Google Patents

Description

The present invention relates to an improvement of a vibration damping device for a railway vehicle.

The railroad vehicle is equipped with a double-acting actuator interposed between the vehicle body and the front and rear trolleys, an acceleration sensor that detects the front-rear acceleration of the vehicle body, and a controller that controls the actuator. In some cases, a vibration damping device for railroad vehicles that suppresses vibration in the left-right direction may be provided.

In such a vibration damping device for railway vehicles, the controller obtains the control force to be generated by the actuator based on the acceleration detected by the acceleration sensor, and causes the actuator to exert a thrust that suppresses the vibration of the vehicle body to vibrate the vehicle body. (See, for example, Patent Document 1).

Japanese Unexamined Patent Publication No. 2013-1304

The actuator in the above-mentioned vibration damping device for railroad vehicles is an electro-hydraulic cylinder, and is expanded and contracted by pressure oil supplied to the cylinder from a pump driven by a motor.

Then, in the vibration damping device for railway vehicles, open-loop control is performed to control the thrust of the actuator by adjusting the pressure in the cylinder with a variable relief valve while supplying pressure oil into the cylinder by rotating the pump at a constant speed with a motor. Is going.

However, in open-loop control, the thrust of the actuator may not be controlled as intended due to the influence of disturbance such as flow rate fluctuation caused by expansion and contraction of the actuator due to an external force. Therefore, in order to obtain a higher damping effect, it is necessary to observe the thrust of the actuator and perform closed loop control to feed back the thrust.

Closed-loop control requires detection of actuator thrust. To detect the thrust, it is possible to directly detect the thrust using a load sensor or to use a pressure sensor that detects the pressure in the cylinder, but it is not common to use a load sensor. In any case, since a sensor is required, there is a problem that the system becomes expensive.

Therefore, an object of the present invention is to provide a vibration damping device for railway vehicles, which has a high vibration damping effect and is inexpensive.

The vibration damping device for a railroad vehicle of the present invention includes an actuator that can expand and contract by supplying a working liquid from a pump driven by a motor, an acceleration sensor, and a controller that controls the actuator, and the controller is derived from the current of the motor. It has an estimation unit that estimates the thrust of the actuator, feeds back the estimated thrust estimated by the estimation unit, and determines the deviation between the target thrust of the actuator and the estimated thrust obtained based on the acceleration detected by the acceleration sensor. The actuator is controlled based on the above. In the vibration damping device for railroad vehicles configured in this way, closed loop control that feeds back thrust can be performed without using a sensor that detects the load of the actuator or the pressure in the cylinder when controlling the actuator. Further, the vibration damping device for a railroad vehicle includes an actuator that can expand and contract by supplying a working liquid from a pump driven by a motor, an acceleration sensor, and a controller that controls the actuator, and the controller is the acceleration sensor. There a target thrust calculating section for obtaining the target force of the actuator is determined based on the acceleration detected, the estimation unit for estimating the thrust of the actuator from the current of the motor, the estimation unit based on the frequency of the target thrust and a correcting unit for correcting the estimated thrust, the correction unit may control the actuator based feeds back the estimated thrust corrected to said estimated thrust and the target thrust by.

Further, when the estimation unit in the vibration damping device for a railroad vehicle determines the polarity which is the expansion / contraction direction of the actuator based on the target thrust of the actuator, the actuator even if the rotation direction of the motor is only one direction. Thrust can be estimated. Therefore, a high-response actuator can be used, and a higher vibration damping effect can be obtained.

The actuator in the vibration damping device for a railroad vehicle may have a switching valve for switching the expansion / contraction direction, and the estimation unit may determine the polarity which is the expansion / contraction direction of the actuator based on the operating state of the switching valve. Also in this case, since it is possible to use a motor that is rotationally driven by the actuator in only one direction, it is possible to use a highly responsive actuator, and a higher vibration damping effect can be obtained.

Further, the vibration damping device for a railway vehicle may include a correction unit that corrects the estimated thrust based on the frequency of the target thrust of the actuator. When the vibration damping device for rolling stock is configured in this way, even if a gear pump is used as the pump, the thrust of the actuator follows the target thrust in the entire frequency range of the target thrust, further improving the riding comfort in the vehicle. can.

Further, the vibration damping device for railroad vehicles is a filter having a characteristic that the controller has a proportional path in which the deviation between the target thrust and the estimated thrust is input, and the correction unit has a characteristic that the gain increases as the frequency of the target thrust of the actuator increases. It may be configured to process the estimated thrust to correct the estimated thrust. When the vibration damping device for rolling stock is configured in this way, it is possible to tune two gains, the gain when correcting the thrust with the correction unit and the proportional gain, and the thrust of the actuator can be adjusted in the entire frequency range of the target thrust. It is possible to more effectively improve the riding comfort in the vehicle by following the target thrust with high accuracy.

In addition, when the correction unit is configured to perform low-pass filtering of the estimated thrust, the vibration damping device for rolling stock can remove the vibration component due to the influence of the structure of the pump from the estimated thrust, and the thrust of the actuator is vibrating. It can be controlled so that it does not become, and the riding comfort can be improved.

Then, the actuator in the vibration damping device for railway vehicles communicates the cylinder, the piston, the rod, the tank, the pump that supplies the working liquid to the rod side chamber, the motor that drives the pump, and the rod side chamber and the piston side chamber. The first on-off valve provided in the middle of the first passage, the second on-off valve provided in the middle of the second passage connecting the piston side chamber and the tank, the discharge passage connecting the rod side chamber and the tank, and the discharge. A variable relief valve provided in the middle of the passage that allows the valve opening pressure to be changed, a rectifying passage that allows only the flow of working liquid from the piston side chamber to the rod side chamber, and a rectifying passage that allows only the flow of working liquid from the tank to the piston side chamber. It may be provided with a suction passage. In the vibration damping device for railway vehicles configured in this way, the actuator functions as a skyhook semi-active damper even when the motor is stopped, so that the vibration damping effect is not lost even when the motor is stopped.

According to the vibration damping device for a railway vehicle of the present invention, a high vibration damping effect can be obtained and the system can be inexpensive in damping the vehicle body of the railway vehicle.

It is a cross section of a railway vehicle equipped with a vibration damping device for a railway vehicle according to the first embodiment. It is a detailed view of an example of an actuator. It is a control block diagram of the controller in the vibration damping device for a railroad vehicle of the first embodiment. It is a figure which showed the relationship between the torque of a motor and the actual thrust of an actuator. It is a control block diagram in one modification of the controller of the vibration damping device for a railroad vehicle of the first embodiment. It is a control block diagram in another modification of the controller of the vibration damping device for a railroad vehicle of the first embodiment. It is a control block diagram of the controller in the vibration damping device for a railroad vehicle of the second embodiment. It is a figure which showed the frequency characteristic of the filter in a correction part.

<First Embodiment>
The railroad vehicle vibration damping device V1 of the first embodiment is used as a vibration damping device for the vehicle body B of the railroad vehicle, and is operated from a pump 12 driven by a motor 15 as shown in FIGS. 1 and 2. It is configured to include an actuator A that can be expanded and contracted by supplying a liquid, and a controller C1 that controls the actuator A.

Specifically, in the case of a railroad vehicle, the actuator A is connected to a pin P hanging below the vehicle body B, and is interposed in parallel between the vehicle body B and the bogie T in pairs. The bogie T holds the wheels W rotatably, and a suspension spring S called a pillow spring is interposed between the bogie B and the bogie T, and the bogie B is elastically supported. Lateral movement of the vehicle body B with respect to the bogie T is permitted.

Then, these actuators A are basically adapted to suppress vibration in the horizontal and lateral directions with respect to the vehicle traveling direction of the vehicle body B by active control by the controller C1.

The controller C1 obtains the target thrust Fref to be generated by the actuator A based on the acceleration α in the horizontal and lateral directions with respect to the vehicle traveling direction of the vehicle body B detected by the acceleration sensor 40, and each actuator A has the target thrust Fref. Gives a command to generate thrust. In this way, the railroad vehicle vibration damping device V1 causes the actuator A to exert a target thrust Fref to suppress the lateral vibration of the vehicle body B.

Next, a specific configuration of the actuator A will be described. It should be noted that one actuator A may be provided with respect to the carriage T, or a plurality of actuators A may be provided. When a plurality of actuators A are provided, the controller C1 may control all the actuators A, or the controller C1 may be provided for each actuator A.

As shown in FIG. 2, the actuator A includes a cylinder 2 connected to one of the vehicle body B and the carriage T of the railway vehicle, a piston 3 slidably inserted into the cylinder 2, and the inside of the cylinder 2. The rod 4 which is inserted into the vehicle body B and the other side of the carriage T and the piston 3 and the tank 7 which stores the working liquid, and the working liquid can be sucked up from the tank 7 and supplied to the rod side chamber 5. It includes a pump 12, a motor 15 for driving the pump 12, and a hydraulic circuit HC for switching expansion and contraction of the actuator A and controlling thrust, and is configured as a single-rod type actuator.

Further, in this example, the rod side chamber 5 and the piston side chamber 6 are filled with hydraulic oil as a hydraulic liquid, and the tank 7 is filled with a gas in addition to the hydraulic oil. It is not necessary to pressurize the inside of the tank 7 by compressing and filling the gas. Further, as the hydraulic liquid, other liquids other than the hydraulic oil may be used.

The hydraulic circuit HC is provided in a first on-off valve 9 as a switching valve provided in a first passage 8 that communicates the rod side chamber 5 and the piston side chamber 6, and a second passage 10 that communicates the piston side chamber 6 and the tank 7. It includes a second on-off valve 11 as a switching valve provided, and a variable relief valve 22 provided in the discharge passage 21 connecting the rod side chamber 5 and the tank 7 and capable of changing the valve opening pressure.

Then, basically, when the first on-off valve 9 makes the first passage 8 communicate, the second on-off valve 11 is closed and the pump 12 is driven, the actuator A is extended and the second on-off valve 11 is the second. When the passage 10 is in a communicating state, the first on-off valve 9 is closed, and the pump 12 is driven, the actuator A contracts.

Hereinafter, each part of the actuator A will be described in detail. The cylinder 2 has a tubular shape, the right end in FIG. 2 is closed by a lid 13, and an annular rod guide 14 is attached to the left end in FIG. Further, a rod 4 that is movably inserted into the cylinder 2 is slidably inserted into the rod guide 14. One end of the rod 4 projects out of the cylinder 2, and the other end of the rod 4 is connected to a piston 3 that is slidably inserted into the cylinder 2.

The outer circumference of the rod guide 14 and the cylinder 2 are sealed by a sealing member (not shown), whereby the inside of the cylinder 2 is maintained in a sealed state. The rod side chamber 5 and the piston side chamber 6 partitioned by the piston 3 in the cylinder 2 are filled with hydraulic oil as described above.

Further, in the case of this actuator A, the cross-sectional area of the rod 4 is halved from the cross-sectional area of the piston 3, and the pressure receiving area on the rod side chamber 5 side of the piston 3 is halved from the pressure receiving area on the piston side chamber 6 side. It has become like. Therefore, if the pressure of the rod side chamber 5 is the same during the extension operation and the contraction operation, the thrust generated in both expansion and contraction becomes equal, and the amount of hydraulic oil with respect to the displacement amount of the actuator A also becomes the same on both sides of expansion and contraction.

Specifically, when the actuator A is extended, the rod side chamber 5 and the piston side chamber 6 are in communication with each other. Then, the pressures in the rod side chamber 5 and the piston side chamber 6 become equal, and the actuator A generates a thrust obtained by multiplying the pressure receiving area difference between the rod side chamber 5 side and the piston side chamber 6 side in the piston 3 by the pressure. On the contrary, when the actuator A is contracted, the rod side chamber 5 and the piston side chamber 6 are cut off from each other, and the piston side chamber 6 is made to communicate with the tank 7. Then, the actuator A generates a thrust obtained by multiplying the pressure in the rod side chamber 5 by the pressure receiving area on the rod side chamber 5 side of the piston 3.

In short, the generated thrust of the actuator A is a value obtained by multiplying half of the cross-sectional area of the piston 3 by the pressure of the rod side chamber 5 in both expansion and contraction. Therefore, when controlling the thrust of the actuator A, the pressure of the rod side chamber 5 may be controlled for both the extension operation and the contraction operation. Further, in the actuator A of this example, since the pressure receiving area on the rod side chamber 5 side of the piston 3 is set to half of the pressure receiving area on the piston side chamber 6 side, the extension side when the same thrust is generated on both sides of expansion and contraction. Since the pressure of the rod side chamber 5 is the same on the contraction side, the control is simplified. In addition, since the amount of hydraulic oil is the same with respect to the amount of displacement, there is an advantage that the responsiveness is the same on both sides of expansion and contraction. Even if the pressure receiving area on the rod side chamber 5 side of the piston 3 is not set to half of the pressure receiving area on the piston side chamber 6, the thrust of the actuator A on both sides of expansion and contraction can be controlled by the pressure of the rod side chamber 5. Does not change.

Returning, the lid 13 that closes the left end of the rod 4 in FIG. 2 and the right end of the cylinder 2 is provided with a mounting portion (not shown), and this actuator A is inserted between the vehicle body B and the bogie T in the railway vehicle. It can be dressed up.

The rod side chamber 5 and the piston side chamber 6 are communicated with each other by a first passage 8, and a first on-off valve 9 is provided in the middle of the first passage 8. The first passage 8 communicates the rod side chamber 5 and the piston side chamber 6 outside the cylinder 2, but may be provided in the piston 3.

The first on-off valve 9 is an electromagnetic on-off valve, and has a communication position in which the first passage 8 is opened to communicate the rod side chamber 5 and the piston side chamber 6, and the first passage 8 is blocked to form the rod side chamber 5. It has a blocking position that cuts off communication with the piston side chamber 6. The first on-off valve 9 takes a communication position when the power is turned on and a shutoff position when the power is off.

Subsequently, the piston side chamber 6 and the tank 7 are communicated with each other by a second passage 10, and a second on-off valve 11 is provided in the middle of the second passage 10. The second on-off valve 11 is an electromagnetic on-off valve, and has a communication position in which the second passage 10 is opened to communicate the piston side chamber 6 and the tank 7, and the second passage 10 is blocked to block the piston side chamber 6 and the tank. It has a blocking position that cuts off communication with 7. The second on-off valve 11 takes a communication position when the power is turned on and a shutoff position when the power is off.

The pump 12 is a gear pump that is driven by a motor 15 that is controlled by the controller C1 and rotates at a predetermined rotation speed, and discharges hydraulic oil in only one direction. The discharge port of the pump 12 is communicated with the rod side chamber 5 by the supply passage 16, and the suction port is connected to the tank 7. When the pump 12 is driven by the motor 15, hydraulic oil is sucked from the tank 7 and the rod is sucked. The hydraulic oil is supplied to the side chamber 5.

As described above, the pump 12 is controlled to rotate at a constant rotation speed, and only discharges hydraulic oil in one direction, and there is no switching operation in the rotation direction. Therefore, there is a problem that the discharge amount changes at the time of rotation switching. Is none. Further, since the rotation direction of the pump 12 is always the same, the motor 15 which is the drive source for driving the pump 12 is not required to have high responsiveness to the rotation switching, and the motor 15 is also inexpensive. Can be used. A check valve 17 is provided in the middle of the supply passage 16 to prevent the backflow of hydraulic oil from the rod side chamber 5 to the pump 12. The motor 15 is driven by receiving electric power from an inverter circuit (not shown) controlled by the controller C1.

Further, in the hydraulic circuit HC of this example, as described above, the discharge passage 21 connecting the rod side chamber 5 and the tank 7 and the variable relief valve 22 provided in the middle of the discharge passage 21 can change the valve opening pressure. It has.

In this example, the variable relief valve 22 is a proportional electromagnetic relief valve, and the valve opening pressure can be adjusted according to the supplied current. When the current is maximized, the valve opening pressure is minimized and the current supply is supplied. Otherwise, the valve opening pressure is maximized.

When the discharge passage 21 and the variable relief valve 22 are provided in this way, the pressure in the rod side chamber 5 can be adjusted to the valve opening pressure of the variable relief valve 22 when the actuator A is expanded and contracted, and the thrust of the actuator A can be adjusted. It can be controlled by the current supplied to the variable relief valve 22. When the discharge passage 21 and the variable relief valve 22 are provided, the sensors required for adjusting the thrust of the actuator A become unnecessary, and the motor 15 does not need to be highly controlled for adjusting the discharge flow rate of the pump 12. .. Therefore, the vibration damping device V1 for railway vehicles becomes inexpensive, and a robust system can be constructed in terms of both hardware and software.

When the first on-off valve 9 is opened and the second on-off valve 11 is closed, or when the first on-off valve 9 is closed and the second on-off valve 11 is opened, vibration input from an external force is applied regardless of the driving condition of the pump 12. On the other hand, the actuator A can exert a damping force only in either extension or contraction. Therefore, for example, when the direction in which the damping force is exerted is the direction in which the vehicle body B is vibrated by the vibration of the bogie T of the railway vehicle, the actuator A is used as a one-sided damper so as not to exert the damping force in such a direction. Can work with. Therefore, since this actuator A can easily realize semi-active control based on Carnop's skyhook theory, it can also function as a semi-active damper.

If a proportional electromagnetic relief valve that changes the valve opening pressure proportionally with the current applied to the variable relief valve 22 is used, the valve opening pressure can be easily controlled, but if the variable relief valve can adjust the valve opening pressure, it is proportional. It is not limited to the electromagnetic relief valve.

In the variable relief valve 22, regardless of the open / closed state of the first on-off valve 9 and the second on-off valve 11, the actuator A receives an excessive input in the expansion / contraction direction, and the pressure in the rod side chamber 5 exceeds the valve opening pressure. When the state is reached, the discharge passage 21 is opened. As described above, when the pressure in the rod side chamber 5 becomes equal to or higher than the valve opening pressure, the variable relief valve 22 discharges the pressure in the rod side chamber 5 to the tank 7 to prevent the pressure in the cylinder 2 from becoming excessive. To protect the entire system of actuator A. Therefore, if the discharge passage 21 and the variable relief valve 22 are provided, the system can be protected.

In addition to the above-described configuration, the hydraulic circuit HC in the actuator A of this example has a rectifying passage 18 that allows only the flow of hydraulic oil from the piston side chamber 6 to the rod side chamber 5, and a rectifying passage 18 that allows only the flow of hydraulic oil from the tank 7 to the piston side chamber 6. It is provided with a suction passage 19 that allows only the flow of hydraulic oil toward. Therefore, when the actuator A expands and contracts while the first on-off valve 9 and the second on-off valve 11 are closed, hydraulic oil is pushed out from the inside of the cylinder 2. Then, since the variable relief valve 22 gives resistance to the flow of the hydraulic oil discharged from the cylinder 2, the actuator A of this example is in a state where the first on-off valve 9 and the second on-off valve 11 are closed. Functions as a uniflow type damper.

More specifically, the rectifying passage 18 communicates the piston side chamber 6 and the rod side chamber 5, and a check valve 18a is provided in the middle, allowing only the flow of hydraulic oil from the piston side chamber 6 to the rod side chamber 5. It is set as a one-way passage. Further, the suction passage 19 communicates the tank 7 with the piston side chamber 6, and is provided with a check valve 19a in the middle, and is a one-way passage that allows only the flow of hydraulic oil from the tank 7 to the piston side chamber 6. Is set to. The rectifying passage 18 can be integrated into the first passage 8 if the shutoff position of the first on-off valve 9 is a check valve, and the suction passage 19 is also the first if the shut-off position of the second on-off valve 11 is a check valve. It can be integrated into two passages 10.

In the actuator A configured in this way, even if both the first on-off valve 9 and the second on-off valve 11 take the shutoff position, the rod side chamber 5 and the piston side chamber 6 are provided in the rectifying passage 18, the suction passage 19, and the discharge passage 21. And the tank 7 is communicated with a string of beads. Further, the rectifying passage 18, the suction passage 19, and the discharge passage 21 are set as one-way passages. Therefore, when the actuator A expands and contracts due to an external force, the hydraulic oil is always discharged from the cylinder 2 and returned to the tank 7 through the discharge passage 21, and the hydraulic oil that is insufficient in the cylinder 2 is discharged from the tank 7 through the suction passage 19 to the cylinder. It is supplied into 2. Since the variable relief valve 22 acts as a resistance to the flow of hydraulic oil and adjusts the pressure in the cylinder 2 to the valve opening pressure, the actuator A functions as a passive uniflow type damper.

Further, when the actuator A fails to energize each device, the first on-off valve 9 and the second on-off valve 11 each take a shutoff position, and the variable relief valve 22 has a maximum valve opening pressure. Functions as a fixed pressure control valve. Therefore, at the time of such a failure, the actuator A automatically functions as a passive damper.

Subsequently, when the actuator A is made to exert a thrust in a desired extension direction, the controller C1 basically rotates the motor 15 to rotate and drive the pump 12 at a predetermined rotation speed, and hydraulic oil is introduced into the cylinder 2. Supply. Then, the first on-off valve 9 is set to the communication position, and the second on-off valve 11 is set to the shutoff position. In this way, the rod side chamber 5 and the piston side chamber 6 are placed in a communicating state, hydraulic oil is supplied to both of them from the pump 12, the piston 3 is pushed to the left in FIG. 2, and the actuator A exerts thrust in the extension direction. Demonstrate. When the pressure in the rod side chamber 5 and the piston side chamber 6 exceeds the valve opening pressure of the variable relief valve 22, the variable relief valve 22 opens and the hydraulic oil is discharged to the tank 7 through the discharge passage 21. Therefore, the pressure in the rod side chamber 5 and the piston side chamber 6 is controlled by the valve opening pressure of the variable relief valve 22 determined by the current applied to the variable relief valve 22. Then, the actuator A is in the extension direction of the value obtained by multiplying the pressure receiving area difference between the piston side chamber 6 side and the rod side chamber 5 side of the piston 3 by the pressure in the rod side chamber 5 and the piston side chamber 6 controlled by the variable relief valve 22. Demonstrate thrust.

On the other hand, when the actuator A exerts a thrust in a desired contraction direction, the controller C1 rotates the motor 15 to supply hydraulic oil from the pump 12 into the rod side chamber 5 while pressing the first on-off valve 9. The shutoff position is set, and the second on-off valve 11 is set to the communication position. In this way, the piston side chamber 6 and the tank 7 are kept in communication with each other, and the hydraulic oil is supplied from the pump 12 to the rod side chamber 5, so that the piston 3 is pushed to the right in FIG. 2 and the actuator A is in the contraction direction. Demonstrate thrust. Then, as described above, when the current of the variable relief valve 22 is adjusted, the actuator A contracts by multiplying the pressure receiving area on the rod side chamber 5 side of the piston 3 by the pressure in the rod side chamber 5 controlled by the variable relief valve 22. Demonstrate thrust in the direction.

Here, since the gear that pushes out the hydraulic oil in the pump 12 receives resistance corresponding to the pressure in the rod side chamber 5, the motor 15 that rotationally drives the pump 12 outputs a torque corresponding to the pressure in the rod side chamber 5. .. That is, the torque output by the motor 15 is proportional to the pressure in the rod side chamber 5, and if the output torque of the motor 15 is known, the pressure in the rod side chamber 5 can be estimated. Then, as described above, the actuator A exerts a thrust according to the pressure in the rod side chamber 5 regardless of whether it expands or contracts. Therefore, if the output torque of the motor 15 is known, the thrust exerted by the actuator A is exerted. Can be estimated.

Further, as can be understood from the above, the first on-off valve 9 and the second on-off valve 11 function as switching valves for switching the expansion / contraction direction when the actuator A exerts thrust.

The actuator A of this example not only functions as an actuator, but can also function as a damper only by opening and closing the first on-off valve 9 and the second on-off valve 11 regardless of the driving condition of the motor 15. Further, when the actuator A is switched from the actuator to the damper, it does not involve a troublesome and steep switching operation between the first on-off valve 9 and the second on-off valve 11, so that a system with high responsiveness and reliability can be provided.

Further, since the actuator A of this example is set to the single rod type, it is easier to secure the stroke length as compared with the double rod type actuator, and the total length of the actuator is shortened, so that the actuator can be used for railway vehicles. The mountability is improved.

The hydraulic oil supply from the pump 12 and the flow of the hydraulic oil due to the expansion / contraction operation in the actuator A of this example pass through the rod side chamber 5 and the piston side chamber 6 in this order, and finally return to the tank 7. Therefore, even if gas is mixed in the rod side chamber 5 or the piston side chamber 6, it is discharged to the tank 7 by the expansion / contraction operation of the actuator A, so that deterioration of the responsiveness of thrust generation can be prevented. Therefore, when manufacturing the actuator A, the troublesome assembly in oil and the assembly in a vacuum environment are not required, and the advanced degassing of hydraulic oil is not required, so that the productivity is improved and the manufacturing cost is reduced. can. Further, even if gas is mixed in the rod side chamber 5 or the piston side chamber 6, the gas is autonomously discharged to the tank 7 by the expansion / contraction operation of the actuator A, so that it is necessary to frequently perform maintenance for performance recovery. This eliminates the labor and cost burden on maintenance.

The configuration of the actuator A is not limited to the above, and for example, a switching valve that allows the pump to communicate with the cylinder by selecting either the extension side chamber or the compression side chamber of the cylinder is provided between the cylinder and the pump. It is also possible to adopt a configuration such as providing. Even with such a configuration, since the pump receives resistance from the pressure in the room where the hydraulic oil is supplied, the thrust of the actuator A can be estimated from the output torque of the motor that drives the pump.

Subsequently, as shown in FIG. 3, the controller C1 of this example has a target thrust Fref to be output by the actuator A based on the acceleration α in the horizontal and lateral directions with respect to the vehicle traveling direction of the vehicle body B detected by the acceleration sensor 40. The target thrust calculation unit 41, the estimation unit 42 that estimates the thrust of the actuator A from the current of the motor 15, and the command voltage according to the current given to the variable relief valve 22 from the deviation ε between the target thrust Fref and the estimated thrust Fm. The relief valve control unit 43 that obtains V, the motor control unit 44 that drives and controls the motor 15 at a predetermined rotation speed, the relief valve drive unit 45 that drives the variable relief valve 22 based on the command voltage V, and the first opening / closing. It includes an on-off valve drive unit 46 that drives and controls the valve 9 and the second on-off valve 11.

The target thrust calculation unit 41 processes the acceleration with a bandpass filter that removes steady acceleration, drift components and noise during curve traveling included in the acceleration α detected by the acceleration sensor 40, and the target thrust that the actuator A should exert. Find the Fref. In this example, the target thrust calculation unit 41 is an H∞ controller, and obtains a target thrust Fref that indicates a thrust to be output by the actuator A in order to suppress vibration of the vehicle body B from the acceleration α. The target thrust Fref is given a positive or negative sign depending on the direction, and the sign indicates the direction of the thrust to be output to the actuator A.

The motor control unit 44 monitors the rotation speed of the motor 15, feeds back the rotation speed of the motor 15 (speed feedback), and controls the motor 15 so as to drive the pump 12 to rotate at the predetermined rotation speed described above. do. More specifically, the motor control unit 44 generates a current command given to the motor 15 from the deviation between the target rotation speed of the motor 15 for rotating the pump 12 at a predetermined rotation speed and the actual rotation speed of the motor 15. To control the motor 15. In this way, the motor control unit 44 controls the motor 15 so that the rotation speed of the motor 15 becomes the target rotation speed.

The estimation unit 42 estimates the thrust exerted by the actuator A and obtains the estimated thrust Fm. Specifically, first, the estimation unit 42 detects the current flowing through the motor 15 and estimates the magnitude of the thrust of the actuator A from this current. As described above, since the pump 12 receives the resistance of the pressure in the rod side chamber 5, the torque of the motor 15 and the thrust of the actuator A are substantially proportional to each other. The torque of the motor 15 is obtained from the current flowing through the motor 15 from the TI characteristics of the motor 15. Therefore, the magnitude of the thrust of the actuator A can be obtained from the current flowing through the motor 15. In this example, the torque of the motor 15 and the actual thrust of the actuator A are measured, and as shown in FIG. 4, the relationship between the torque of the motor 15 and the actual thrust of the actuator A is grasped in advance. Then, if this relationship is mathematically expressed or mapped, the torque can be obtained from the current flowing through the motor 15, and the magnitude of the thrust of the actuator A can be easily estimated from the obtained torque. In formulating, for example, an approximate expression may be obtained by using the least squares method or the like, and this approximate expression may be used as a mathematical expression. In FIG. 4, the thrust of the actuator A is not exerted unless the torque of the motor 15 is t1 or more because of the friction between the actuator A, the pump 12, and the motor 15. In this way, the magnitude of the thrust exerted by the actuator A can be obtained from the current flowing through the motor 15, but it is necessary to determine whether the direction of the thrust of the actuator A is the extension direction or the contraction direction.

Therefore, the estimation unit 42 determines from the sign of the target thrust Fref obtained by the target thrust calculation unit 41 whether the direction in which the actuator A should exert the thrust is the extension direction or the contraction direction. That is, the polarity is determined. In the target thrust Fref, the numerical value indicates the magnitude of the thrust of the actuator A and the reference numeral indicates the polarity which is the direction of the thrust of the actuator A, and the estimation unit 42 determines the polarity using the reference numeral. In this example, when the actuator A exerts a thrust in the extension direction, the target thrust Fref takes a positive value, and conversely, when the actuator A exerts a thrust in the contraction direction, the target thrust Fref takes a negative value. It is set as. In this way, the estimation unit 42 obtains the magnitude of the thrust of the actuator A from the current of the motor 15, determines the polarity from the sign of the target thrust Ref, estimates the thrust, and obtains the estimated thrust Fm. Therefore, for example, when the magnitude of the thrust of the actuator A obtained from the current of the motor 15 is a, the estimation unit 42 estimates that the polarity is positive and indicates the extension direction, and estimates that the estimated thrust Fm is + a, and the polarity is positive. Is negative and indicates the contraction direction, and the estimated thrust Fm is estimated to be -a. The target thrust Fref is set to take a negative value when the actuator A exerts a thrust in the extension direction, and the target thrust Fref is set to take a positive value when the actuator A exerts a thrust in the contraction direction. May be good.

The relief valve control unit 43 obtains the command voltage V from the deviation ε between the target thrust Fref and the estimated thrust Fm obtained by the target thrust calculation unit 41. In this example, the relief valve control unit 43 is a proportional compensator, and is a deviation calculation unit 43a for obtaining the deviation ε between the target thrust Fref and the estimated thrust Fm, and an absolute value processing unit 43b for processing the deviation ε in absolute value. And the gain multiplication unit 43c that multiplies the deviation | ε | processed by the absolute value by the proportional gain K, and the command voltage that obtains the command voltage V given to the relief valve drive unit 45 from the value | ε | K obtained by the gain multiplication unit 43c. It is provided with a calculation unit 43d. The command voltage calculation unit 43d has in advance the relationship between the thrust of the actuator A and the voltage to be applied to the relief valve drive unit 45 for realizing this thrust as a map or a mathematical formula. Therefore, the command voltage calculation unit 43d uses the above-mentioned map or mathematical formula to obtain the command voltage V to the relief valve drive unit 45 with the value | ε | K as a parameter.

The relief valve drive unit 45 includes a driver circuit for driving the variable relief valve 22, receives an input of a command voltage V from the relief valve control unit 43, and changes the current according to the command voltage V to the variable relief valve 22. It is supplied to the relief valve 22. In this way, the relief valve drive unit 45 adjusts the current supplied to the variable relief valve 22 according to the command voltage V to control the valve opening pressure of the variable relief valve 22.

The on-off valve drive unit 46 determines the polarity in the expansion / contraction direction of the actuator A from the code of the target thrust Fref obtained by the target thrust calculation unit 41, and drives and controls the first on-off valve 9 and the second on-off valve 11. When the polarity of the actuator A indicated by the target thrust Fref is in the extension direction, the on-off valve drive unit 46 drives the first on-off valve 9 and the second on-off valve 11 to set the first on-off valve 9 in the communication position. (Ii) The on-off valve 11 is set to the shutoff position. On the other hand, when the polarity of the actuator A indicated by the target thrust Fref is in the contraction direction, the on-off valve drive unit 46 drives the first on-off valve 9 and the second on-off valve 11 to shut off the first on-off valve 9. The second on-off valve 11 is set to the communication position.

The target thrust calculation unit 41, the estimation unit 42, and the relief valve control unit 43 in the controller C1 are not shown as hardware resources, but specifically, for example, A for capturing a signal output by the acceleration sensor 40. Based on the / D converter, a storage device such as a ROM (Read Only Memory) that stores a program used for processing required to capture the output value of the acceleration sensor 40 and control the actuator A, and the program. It suffices to be configured to include a computing device such as a CPU (Central Processing Unit) that executes the processing and a storage device such as a RAM (Random Access Memory) that provides a storage area to the CPU. It can be realized by execution.

As described above, the railroad vehicle vibration damping device V1 of the first embodiment has an actuator A that can be expanded and contracted by supplying hydraulic oil from a pump 12 driven by a motor 15 and an actuator A that can be expanded and contracted by supplying hydraulic oil from the current of the motor 15. A controller C1 is provided which has an estimation unit 42 for estimating the thrust of the above and feeds back the estimated thrust Fm estimated by the estimation unit 42 to control the actuator A.

Therefore, the controller C1 obtains the target thrust Fref based on the acceleration α detected by the acceleration sensor 40, estimates the thrust of the actuator A based on the current flowing through the motor 15, feeds back the estimated thrust Fm, and performs closed loop control. Can control the actuator A.

As described above, in the vibration damping device V1 for a railroad vehicle of the present invention, when controlling the actuator A, the estimated thrust Fm is fed back without using a sensor that detects the load of the actuator A and the pressure in the cylinder 2. Closed loop control can be performed.

Therefore, according to the vibration damping device V1 for a railroad vehicle of the present invention, the thrust of the actuator A can be made to follow the target thrust Fref even if there is an input of disturbance, so that a high vibration damping effect can be obtained. Further, according to the vibration damping device V1 for a railroad vehicle of the present invention, it is not necessary to install a sensor for detecting the load of the actuator A and the pressure in the cylinder 2 in the closed loop control, so that the system becomes inexpensive. From the above, according to the vibration damping device V1 for a railway vehicle of the present invention, a high vibration damping effect can be obtained and the system can be inexpensive in damping the vehicle body B of the railway vehicle. Further, according to the vibration damping device V1 for a railroad vehicle, the tuning work is troublesome because the accuracy of the valve opening pressure of the variable relief valve 22 is required in the open loop control, but the estimated thrust of the actuator A in the closed loop control. Since the valve opening pressure of the variable relief valve 22 is automatically adjusted by the feedback, the tuning work of the valve opening pressure of the variable relief valve 22 becomes very easy.

Further, since the estimation unit 42 determines the polarity of the actuator A based on the target thrust Fref of the actuator A, the thrust of the actuator A can be estimated even when the rotation direction of the motor 15 is only one direction. When the structure of the actuator A is such that a bidirectional discharge type pump is provided in the middle of the passage communicating the extension side chamber and the compression side chamber of the cylinder, the rotation direction of the motor is switched to expand and contract the actuator A. It will be switched. In the case of the actuator A having such a configuration, the estimation unit 42 can also determine the polarity from the current flowing through the motor, so that the estimated thrust Fm can be obtained only from the current. The actuator A may be configured in this way, and the thrust of the actuator A may be estimated by the estimation unit 42 only from the current flowing through the motor to obtain the estimated thrust Fm. However, when the motor rotates in both directions, the rotation direction of the motor Due to the influence of inertia when the actuator A is switched, it becomes difficult for the thrust to follow the target thrust Fref when the actuator A is switched in the expansion / contraction direction. Therefore, when the estimation unit 42 determines the polarity of the actuator A based on the target thrust Fref of the actuator A, the motor 15 that is rotationally driven by the actuator A in only one direction can be used, and the expansion / contraction direction of the actuator A can be switched. In the meantime, the thrust easily follows the target thrust Fref. From the above, according to the vibration damping device V1 for a railroad vehicle of this example, in which the estimation unit 42 determines the polarity of the actuator A in the expansion / contraction direction based on the target thrust Fref of the actuator A, the actuator A with high response can be used. Therefore, a higher damping effect can be obtained.

In the above-mentioned place, the target thrust Fref obtained by the target thrust calculation unit 41 is used for the polarity determination, but the estimation unit 42 operates the first on-off valve 9 and the second on-off valve 11 as switching valves. Situation, in this example, the polarity may be determined from the opening / closing status of the first on-off valve 9 and the second on-off valve 11. When the actuator A is extended, the on-off valve drive unit 46 supplies a current to the first on-off valve 9 and does not supply a current to the second on-off valve 11. Further, when the actuator A is contracted, the on-off valve drive unit 46 does not supply a current to the first on-off valve 9, but supplies a current to the second on-off valve 11. As described above, in this example, the polarity of the actuator A can be determined from the opening / closing status of the first on-off valve 9 and the second on-off valve 11. Therefore, when determining the polarity of the actuator A, as shown in FIG. 5, the estimation unit 42 monitors the excitation state of the first on-off valve 9 and the second on-off valve 11, which are switching valves, or monitors the operating status of both by an excitation signal. The thrust of the actuator A may be estimated by grasping and determining the polarity. When the first on-off valve 9 and the second on-off valve 11 have means for sensing their own positions, the estimation unit 42 uses the first on-off valve 9 and the first on-off valve 9 which are switching valves from the positions obtained from the means. The polarity may be determined by grasping the operating state of the second on-off valve 11. In short, the estimation unit 42 may determine the polarity based on the operating condition of the switching valve. When the vibration damping device V1 for a railway vehicle is configured in this way, the expansion / contraction direction of the actuator A can be accurately grasped, so that the thrust of the actuator A can be accurately estimated. Further, also in this case, since the motor 15 which is rotationally driven by the actuator A in only one direction can be used, the actuator A having a high response can be used, and a higher vibration damping effect can be obtained.

As shown in FIG. 6, the relief valve control unit 43 is estimated to be an absolute value processing unit 43e that processes the target thrust Fref in absolute value instead of the absolute value processing unit 43b that processes the deviation ε in absolute value. It may be configured to include an absolute value processing unit 43f that processes the thrust Fm by an absolute value. The deviation calculation unit 43a calculates the deviation ε between the absolute value-processed target thrust | Fref | and the absolute value-processed estimated thrust | Fm |, and the gain multiplication unit 43c and the command voltage calculation unit 43d subsequently calculate the deviation ε. You just have to do the processing.

Further, as described above, the relief valve control unit 43 includes only the proportional path for multiplying the proportional gain K by the deviation ε and performs only the proportional compensation. However, in addition to the proportional path, the integral path or the integral path is performed. And may have a configuration in which a differential path is added.

Further, the vibration damping device V1 for a railroad vehicle of this example includes a cylinder 2, a piston 3, a rod 4, a tank 7, a pump 12 that supplies hydraulic oil to a rod side chamber 5, and a motor 15 that drives the pump 12. The first on-off valve 9 provided in the first passage 8 that communicates the rod side chamber 5 and the piston side chamber 6, and the second on-off valve 11 provided in the second passage 10 that communicates the piston side chamber 6 and the tank 7. , A variable relief valve 22 provided in the discharge passage 21 connecting the rod side chamber 5 and the tank 7 and capable of changing the valve opening pressure, and a rectifying passage 18 that allows only the flow of hydraulic oil from the piston side chamber 6 to the rod side chamber 5. And a suction passage 19 that allows only the flow of hydraulic oil from the tank 7 to the piston side chamber 6. In the vibration damping device V1 for railway vehicles configured in this way, even if the pump 12 is stopped, the actuator A functions as a skyhook semi-active damper, so that the vibration damping effect is not lost even when the pump 12 is stopped. ..

<Second embodiment>
As shown in FIG. 7, the railroad vehicle vibration damping device V2 of the second embodiment has a different configuration in the controller C2 from the railroad vehicle vibration damping device V1 of the first embodiment. In the railroad vehicle vibration damping device V2 of the second embodiment, the controller C2 corrects the estimated thrust Fm based on the frequency of the target thrust Fref in the configuration of the controller C1 of the first embodiment. Is added.

As described above, the pump 12 of the actuator A in this example is a gear pump. The gear pump is designed so that the two gears rotate while meshing with each other to suck the hydraulic oil from the suction port and discharge it from the discharge port, but the resistance of the meshing of the gears changes due to the backlash. On the other hand, the motor 15 for driving the pump 12 is controlled by the motor control unit 44 so as to rotate at a constant speed at the above-mentioned predetermined rotation speed. Therefore, the pump 12 rotates at a constant speed at a predetermined rotation speed, but the torque itself fluctuates while the motor 15 rotates at a constant speed because the resistance of meshing between the gears changes. This torque fluctuation appears periodically with the rotation of the pump 12.

Therefore, since the current supplied from the motor control unit 44 to the motor 15 also fluctuates, the estimated thrust Fm of the actuator A estimated by the estimation unit 42 also pulsates, so that the obtained estimated thrust Fm also fluctuates in a wavy manner. Therefore, for example, even when the target thrust Fref takes a constant value, the estimated thrust Fm fluctuates, so that the deviation ε also fluctuates. If the value of the proportional gain K is set high, the thrust of the actuator A can easily follow the target thrust Fref, but if this is done, the command voltage V obtained by the influence of the fluctuation of the estimated thrust Fm also undulates greatly. The thrust that the actuator A actually outputs as a waveform also fluctuates greatly. The fluctuation of the thrust of the actuator A gives an extra vibration to the vehicle body B, which may be one of the causes for deteriorating the riding comfort. As described above, although there is a desire to increase the proportional gain K from the viewpoint of controllability, if the value of the proportional gain K is not kept low, the riding comfort will deteriorate. Deterioration of riding comfort becomes remarkable when the target thrust Fref is a low frequency, and does not become a problem so much when the target thrust is a high frequency.

Therefore, the railroad vehicle vibration damping device V2 of the second embodiment includes a correction unit 50 that corrects the estimated thrust Fm. The correction unit 50 includes a filter 50a that filters the estimated thrust Fm obtained by the estimation unit 42, and a gain multiplication unit 50b that multiplies the estimated thrust Fm filtered by the filter 50a by the gain Ky.

As shown in FIG. 8, the filter 50a is a filter having a characteristic that the gain increases as the frequency of the target thrust Fref obtained by the target thrust calculation unit 41 increases. The characteristics of the filter 50a are that the maximum value of the gain is 0 dB and the minimum value is -6 dB. It is set to a characteristic that converges to -6 dB when it asymptotically approaches and becomes a lower frequency. Further, the gain Ky that the gain multiplication unit 50b multiplies the estimated thrust Fm after filtering is set to a constant of 1 or less.

Therefore, when the frequency of the target thrust Fref is less than 10 Hz, the correction unit 50 is corrected so that the numerical value excluding the sign of the estimated thrust Fm becomes smaller according to the frequency. Further, when the frequency of the target thrust Fref is 10 Hz or more, the correction unit 50 outputs a value close to the value obtained by multiplying the estimated thrust Fm by the gain Ky. When the frequency of the target thrust Fref is 10 Hz or less, the numerical values excluding the sign of the estimated thrust Fm are corrected to be small, so that the estimated thrust Fm estimated by the estimation unit 42 fluctuates in a wavy manner due to the influence of the structure of the pump 12. However, the wave height of this fluctuation is corrected to be small.

When the correction unit 50 corrects the estimated thrust Fm in this way, when the frequency of the target thrust Fref is a low frequency of less than 10 Hz, the numerical value excluding the sign of the estimated thrust Fm input to the relief valve control unit 43 is the actual value. Since it is smaller than the value of the thrust of the actuator A, the deviation ε obtained by the deviation calculation unit 43a becomes large. However, the value of the proportional gain K obtained by multiplying the deviation ε by the gain multiplying unit 43c in the relief valve control unit 43 is set to a value smaller than that of the first embodiment in proportion to the increase in the deviation ε. .. Proportional gain so that the thrust of the actuator A follows the target thrust Fref without the command voltage V becoming excessive even if the gain in the filter 50a becomes small and the deviation ε becomes large when the frequency of the target thrust Fref is 10 Hz or less. Both K and gain Ky are tuned.

In this way, when the correction unit 50 is provided, when the frequency of the target thrust Fref is a low frequency of less than 10 Hz, the fluctuation of the estimated thrust Fm after correction is suppressed, so that the fluctuation of the command voltage V is also suppressed, and the actuator The thrust of A can be made to follow the target thrust Fref with high accuracy.

On the other hand, when the frequency of the target thrust Fref is a high frequency of 10 Hz or more, if the correction unit 50 corrects the numerical value excluding the sign of the estimated thrust Fm to a small value and the deviation ε becomes large, the command voltage V becomes excessive. However, when the frequency of the target thrust Fref is a high frequency of 10 Hz or more, the correction unit 50 outputs a value obtained by multiplying the estimated thrust Fm estimated by the estimation unit 42 by the gain Ky. There is not much difference from the corrected estimated thrust Fm. Therefore, the estimated thrust Fm input to the relief valve control unit 43 is a value close to the thrust actually output by the actuator A. When the target thrust Fref is a high frequency, the fluctuation due to the influence of the structure of the pump 12 of the estimated thrust Fm estimated by the estimation unit 42 has little adverse effect on the control in the controller C2. From the above, even if the frequency of the target thrust Fref becomes a high frequency of 10 Hz or more, the command voltage V is prevented from becoming excessive, and the thrust of the actuator A can be made to follow the target thrust Fref with high accuracy.

As described above, since the rolling stock damping device V2 according to the second embodiment includes the correction unit 50 that corrects the estimated thrust Fm based on the frequency of the target thrust Fref, a gear pump is used for the pump 12. Even so, the thrust of the actuator A can be made to follow the target thrust Fref in the entire frequency range of the target thrust Fref, and the riding comfort in the vehicle can be further improved.

Therefore, according to the railroad vehicle vibration damping device V2 in the second embodiment, not only the effect of the railroad vehicle vibration damping device V1 in the first embodiment is exerted, but also the riding comfort in the vehicle is more effective. Can be improved. Further, when the correction unit 50 is configured as described above, the correction unit 50 corrects the estimated thrust Fm without performing the process of causing a time delay, so that the control performance does not deteriorate.

As described above, the relief valve control unit 43 includes only the proportional path for multiplying the proportional gain K by the deviation ε and performs only the proportional compensation. However, in addition to the proportional path, the integral path or the integral path is performed. And may have a configuration in which a differential path is added. Further, the frequency characteristic of the gain in the filter 50a in the correction unit 50 may be a characteristic that the gain increases as the frequency of the target thrust Fref increases, but the minimum and maximum values of the gain and the degree of change according to the frequency of the gain Etc. can be appropriately changed so as to be suitable for vibration suppression of the vehicle body B of the railway vehicle.

Further, since the rolling stock damping device V2 in the second embodiment has a proportional path for multiplying the proportional gain K by the deviation ε, it is proportional to the gain Ky when the correction unit 50 corrects the estimated thrust Fm. It is possible to tune the two gains of the gain K, and the thrust of the actuator A can be made to follow the target thrust Fref with high accuracy in the entire frequency range of the target thrust Fref, so that the riding comfort in the vehicle can be further improved.

The correction unit 50 may be used as a low-pass filter to perform low-pass filter processing in which the estimated thrust Fm estimated by the influence of the structure of the pump 12 fluctuates in a wavy manner. In this way, since the vibration component of the estimated thrust Fm can be removed, it is not necessary to set the proportional gain K small, and the thrust of the actuator A can be controlled so as not to be vibrating, and the riding comfort can be improved.

Although the preferred embodiments of the present invention have been described in detail above, they can be modified, modified, and modified as long as they do not deviate from the claims.

2 ... Cylinder, 3 ... Piston, 4 ... Rod, 5 ... Rod side chamber, 6 ... Piston side chamber, 7 ... Tank, 8 ... First passage, 9 ... First on-off valve (switching valve), 10 ... second passage, 11 ... second on-off valve (switching valve), 12 ... pump, 15 ... motor, 18 ... rectifying passage, 19 ... suction passage, 21 ... discharge passage, 22 ... variable relief valve, 42 ... estimation unit, 50 ... correction unit, A ... actuator, C1, C2 ... controller, V1 , V2 ・ ・ ・ Vibration damping device for railway vehicles

Claims (8)

An actuator that can be expanded and contracted by supplying working liquid from a pump driven by a motor,
Accelerometer and
A controller for controlling the actuator is provided.
The controller
It has an estimation unit that estimates the thrust of the actuator from the current of the motor, feeds back the estimated thrust estimated by the estimation unit, and obtains the target thrust of the actuator based on the acceleration detected by the acceleration sensor and the said. A vibration damping device for railroad vehicles, characterized in that the actuator is controlled based on a deviation of an estimated thrust. An actuator that can be expanded and contracted by supplying working liquid from a pump driven by a motor,
Accelerometer and
A controller for controlling the actuator is provided.
The controller includes a target thrust calculation unit that obtains the target thrust of the actuator obtained based on the acceleration detected by the acceleration sensor, an estimation unit that estimates the thrust of the actuator from the current of the motor, and a frequency of the target thrust. It has a correction unit that corrects the thrust estimated by the estimation unit based on the above, feeds back the estimated thrust corrected by the correction unit, and controls the actuator based on the target thrust and the estimated thrust. A vibration damping device for railway vehicles. The vibration damping device for a railway vehicle according to claim 1 or 2, wherein the estimation unit determines the polarity in the expansion / contraction direction of the actuator based on the target thrust. The actuator has a switching valve for switching the expansion / contraction direction.
The vibration damping device for a railway vehicle according to claim 1 or 2, wherein the estimation unit determines the polarity in the expansion / contraction direction of the actuator based on the operating state of the switching valve. The vibration damping device for a railway vehicle according to claim 1, further comprising a correction unit that corrects the estimated thrust based on the frequency of the target thrust. The controller has a proportional path into which the deviation between the target thrust and the estimated thrust is input.
The rolling stock system according to claim 5, wherein the correction unit processes the estimated thrust with a filter having a characteristic that the gain increases as the frequency of the target thrust increases, and corrects the estimated thrust. Shaking device. The vibration damping device for a railway vehicle according to claim 5, wherein the correction unit performs a low-pass filter process on the estimated thrust. The actuator
Cylinder and
A piston that is slidably inserted into the cylinder,
A rod inserted into the cylinder and connected to the piston,
A rod side chamber and a piston side chamber partitioned by the piston in the cylinder,
With the tank
The pump capable of sucking the working liquid from the tank and supplying the working liquid to the rod side chamber,
The motor that drives the pump and
A first on-off valve provided in a first passage communicating the rod side chamber and the piston side chamber,
A second on-off valve provided in a second passage communicating the piston side chamber and the tank,
A variable relief valve provided in the discharge passage connecting the rod side chamber and the tank,
A rectifying passage that allows only the flow of hydraulic oil from the piston side chamber to the rod side chamber,
The vibration damping device for a railway vehicle according to any one of claims 1 to 7, further comprising a suction passage that allows only the flow of hydraulic oil from the tank to the piston side chamber.

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