WO2018159511A1

WO2018159511A1 - Railroad car vibration damping device

Info

Publication number: WO2018159511A1
Application number: PCT/JP2018/006867
Authority: WO
Inventors: 貴之小川
Original assignee: Ｋｙｂ株式会社
Priority date: 2017-03-03
Filing date: 2018-02-26
Publication date: 2018-09-07
Also published as: JP6924043B2; JP2018144599A

Abstract

This railroad car vibration damping device (V1) comprises: an actuator (A) which can be expanded and contracted by hydraulic fluid supplied from a pump (12) that is driven by a motor (15); and a controller (C1) which has an estimation unit (42) for estimating the thrust of the actuator (A) from the current of the motor (15), and which feeds back the estimation thrust (Fm) estimated by the estimation unit (42) and controls the actuator (A).

Description

Vibration control device for railway vehicles

The present invention relates to an improvement in a railcar vibration damping device.

A rail vehicle includes a double-acting actuator interposed between a vehicle body and front and rear carriages, an acceleration sensor that detects acceleration in the front and rear of the vehicle body, and a controller that controls the actuator. In some cases, a railcar damping device that suppresses vibration in the left-right direction is provided.

In such a railway vehicle vibration damping device, as disclosed in JP2013-1304A, for example, the controller obtains a control force to be generated by the actuator based on the acceleration detected by the acceleration sensor, and the actuator is connected to the vehicle body. The thrust of the vibration is exerted to suppress the vibration of the vehicle body (see, for example, Patent Document 1).

The actuator in the above-described railcar vibration damping device is an electro-hydraulic cylinder, which is extended and contracted by pressure oil supplied to the cylinder from a pump driven by a motor.

In a railcar vibration damping device, an open-loop control that controls the thrust of the actuator by adjusting the pressure in the cylinder with a variable relief valve while supplying the pressure oil into the cylinder by rotating the pump at a constant speed with a motor. Is going.

However, in the open loop control, there is a case where the thrust of the actuator cannot be controlled as intended due to the influence of a disturbance such as a flow rate fluctuation caused by the expansion and contraction of the actuator due to an external force. Therefore, in order to obtain a higher vibration damping effect, it is necessary to perform closed loop control in which the thrust of the actuator is observed and the thrust is fed back.

閉 Closed loop control requires detection of actuator thrust. In order to detect the thrust, a method of directly detecting the thrust using a load sensor or a method using a pressure sensor for detecting the pressure in the cylinder can be considered, 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 an inexpensive railway vehicle vibration damping device that can obtain a high vibration damping effect.

Therefore, the railcar damping device of the present invention includes an actuator that can be expanded and contracted by supplying a working liquid from a pump driven by a motor, and an estimation unit that estimates the thrust of the actuator from the current of the motor. And a controller for controlling the actuator by feeding back the estimated thrust estimated by the estimation unit. In the railcar damping device configured as described above, 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 in controlling the actuator.

FIG. 1 is a cross-sectional view of a railway vehicle on which the railway vehicle vibration damping device according to the first embodiment is mounted. FIG. 2 is a detailed view of an example of the actuator. FIG. 3 is a control block diagram of a controller in the railcar damping device of the first embodiment. FIG. 4 is a diagram showing the relationship between the motor torque and the actual thrust of the actuator. FIG. 5 is a control block diagram in a modified example of the controller of the railcar vibration damping device of the first exemplary embodiment. FIG. 6 is a control block diagram in another modified example of the controller of the railcar vibration damping device of the first exemplary embodiment. FIG. 7 is a control block diagram of a controller in the railcar vibration damping device of the second embodiment. FIG. 8 is a diagram illustrating the frequency characteristics of the filter in the correction unit.

<First embodiment>
The railcar damping device V1 according to the first embodiment is used as a damping device for the vehicle body B of the railcar, and is operated from a pump 12 driven by a motor 15 as shown in FIGS. An actuator A that can be expanded and contracted by supplying liquid and a controller C1 that controls the actuator A are provided.

In detail, in the case of a railway vehicle, the actuator A is connected to a pin P suspended below the vehicle body B, and is interposed between the vehicle body B and the carriage T in parallel. The carriage T rotatably holds the wheel W, and a suspension spring S called a pillow spring is interposed between the vehicle body B and the carriage T, and the vehicle body B is elastically supported. Movement of the vehicle body B in the lateral direction relative to the carriage T is allowed.

These actuators A are basically configured 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 horizontal direction with respect to the vehicle traveling direction of the vehicle body B detected by the acceleration sensor 40, and applies each actuator A according to the target thrust Fref. Gives a command to generate thrust. In this way, the railcar vibration damping device V1 causes the actuator A to exert the target thrust Fref to suppress the lateral vibration of the vehicle body B.

Next, a specific configuration of the actuator A will be described. One actuator A or a plurality of actuators A may be provided for the carriage T. When a plurality of actuators A are provided, all the actuators A may be controlled by the controller C1, or a controller C1 may be provided for each actuator A.

In this example, as shown in FIG. 2, the actuator A includes a cylinder 2 connected to one of a vehicle body B and a carriage T of a railway vehicle, a piston 3 slidably inserted into the cylinder 2, Rod 4 connected to the other of the vehicle body B and the carriage T and the piston 3, a tank 7 for storing the working liquid, and the working liquid can be sucked up from the tank 7 and supplied to the rod side chamber 5. A pump 12, a motor 15 that drives the pump 12, and a hydraulic circuit HC that controls the switching of the expansion and contraction of the actuator A and the thrust are 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 working oil as working liquid, and the tank 7 is filled with gas in addition to working oil. In addition, it is not necessary to compress and fill the inside of the tank 7 with a gas in particular. In addition to the working oil, other liquids may be used as the working liquid.

The hydraulic circuit HC includes a first on-off valve 9 as a switching valve provided in the 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. A second opening / closing valve 11 as a switching valve provided, and a variable relief valve 22 capable of changing the valve opening pressure provided in the discharge passage 21 connecting the rod side chamber 5 and the tank 7 are provided.

Basically, when the first opening / closing valve 9 is in communication with the first passage 8, the second opening / closing valve 11 is closed and the pump 12 is driven, the actuator A extends, and the second opening / closing valve 11 causes the second opening / closing valve 11 to 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 cylindrical 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. 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 protrudes outside the cylinder 2, and the other end in the cylinder 2 is connected to a piston 3 that is slidably inserted into the cylinder 2.

The space between the outer periphery of the rod guide 14 and the cylinder 2 is sealed by a seal 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.

In the case of this actuator A, the rod 4 has a cross-sectional area that is ½ of the cross-sectional area of the piston 3, and the pressure-receiving area of the piston 3 on the rod-side chamber 5 side is ½ of the pressure-receiving area on the piston-side chamber 6 side. It is like that. Therefore, when the pressure in the rod side chamber 5 is the same during the extension operation and during 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 becomes the same in both 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 communication between the rod side chamber 5 and the piston side chamber 6 is cut off and the piston side chamber 6 is connected to 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 of the piston 3 on the rod side chamber 5 side.

In short, the thrust generated by the actuator A is a value obtained by multiplying a half of the cross-sectional area of the piston 3 by the pressure in the rod side chamber 5 in both expansion and contraction. Therefore, when the thrust of the actuator A is controlled, the pressure in the rod side chamber 5 may be controlled for both the extension operation and the contraction operation. Further, in the actuator A of the present example, the pressure receiving area on the rod side chamber 5 side of the piston 3 is set to one half of the pressure receiving area on the piston side chamber 6 side. Since the pressure in the rod side chamber 5 is the same on the contraction side, the control is simplified. In addition, since the amount of hydraulic oil with respect to the amount of displacement is the same, there is an advantage that the responsiveness is the same on both sides of expansion and contraction. In addition, even when the pressure receiving area on the rod side chamber 5 side of the piston 3 is not set to ½ of the pressure receiving area on the piston side chamber 6 side, the thrust on both sides of the actuator A can be controlled by the pressure of the rod side chamber 5. Will not change.

Returning, the lid 4 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 interposed between the vehicle body B and the carriage T in the railway vehicle. Can be disguised.

The rod side chamber 5 and the piston side chamber 6 communicate with each other through a first passage 8, and a first opening / closing 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. The first on-off valve 9 is opened to connect the rod-side chamber 5 and the piston-side chamber 6, and the first on-off passage 8 is shut off to connect to the rod-side chamber 5. And a blocking position for disconnecting communication with the piston side chamber 6. And this 1st on-off valve 9 takes a communicating position at the time of electricity supply, and takes a cutoff position at the time of non-energization.

Subsequently, the piston side chamber 6 and the tank 7 are communicated with each other by a second passage 10, and a second opening / closing valve 11 is provided in the middle of the second passage 10. The second on-off valve 11 is an electromagnetic on-off valve, which opens the second passage 10 to communicate the piston side chamber 6 and the tank 7, and shuts off the second passage 10 to connect the piston side chamber 6 and the tank. 7 and a shut-off position that cuts off communication with 7. And this 2nd on-off valve 11 takes a communicating position at the time of electricity supply, and takes a cutoff position at the time of non-energization.

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

As described above, the pump 12 is controlled to rotate at a constant rotational speed, and only discharges hydraulic oil in one direction, and there is no rotation direction switching operation. There is nothing. Further, since the rotation direction of the pump 12 is always the same direction, even the motor 15 that is a drive source for driving the pump 12 does not require high responsiveness to rotation switching, and the motor 15 is also inexpensive. Can be used. A check valve 17 that prevents the backflow of hydraulic oil from the rod side chamber 5 to the pump 12 is provided in the supply passage 16. The motor 15 is driven by receiving power supply from an inverter circuit (not shown) controlled by the controller C1.

Further, as described above, the hydraulic circuit HC of the present example includes the discharge passage 21 that connects the rod side chamber 5 and the tank 7, and the variable relief valve 22 that can change the valve opening pressure provided in the middle of the discharge passage 21. 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 reaches the maximum, the valve opening pressure is minimized and the supply of current is reduced. Otherwise, the valve opening pressure is maximized.

As described above, when the discharge passage 21 and the variable relief valve 22 are provided, when the actuator A is expanded and contracted, the pressure in the rod side chamber 5 can be adjusted to the valve opening pressure of the variable relief valve 22, and the thrust of the actuator A can be increased. 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, sensors necessary for adjusting the thrust force of the actuator A are not necessary, and it is not necessary to highly control the motor 15 for adjusting the discharge flow rate of the pump 12. . Therefore, the railcar damping device V1 is inexpensive, and a robust system can be constructed in terms of 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 state of the pump 12. On the other hand, the actuator A can exhibit a damping force only in one of expansion and 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 provided with a one-effect damper so that no damping force is generated in such a direction. And can function. Therefore, since this actuator A can easily realize semi-active control based on Karnop's Skyhook theory, it can also function as a semi-active damper.

The use of a proportional electromagnetic relief valve that proportionally changes the valve opening pressure with the current applied to the variable relief valve 22 makes it easy to control the valve opening pressure. However, any variable relief valve that can adjust the valve opening pressure is proportional. It is not limited to an electromagnetic relief valve.

In the variable relief valve 22, regardless of whether the first on-off valve 9 and the second on-off valve 11 are open or closed, the actuator A has 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, the variable relief valve 22 discharges the pressure in the rod side chamber 5 to the tank 7 when the pressure in the rod side chamber 5 becomes equal to or higher than the valve opening pressure, so that the pressure in the cylinder 2 is prevented from becoming excessive. To protect the entire system of the 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 includes 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 the tank 7 to the piston side chamber 6. A suction passage 19 that allows only the flow of hydraulic oil toward the head is provided. 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 cylinder 2. Since the variable relief valve 22 provides resistance to the flow of hydraulic oil discharged from the cylinder 2, the actuator A of the present 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 18 a is provided in the middle, allowing only the flow of hydraulic oil from the piston side chamber 6 toward the rod side chamber 5. It is set as a one-way passage. Further, the suction passage 19 communicates between the tank 7 and the piston side chamber 6, and a check valve 19 a is provided in the middle to allow only the flow of hydraulic oil from the tank 7 toward the piston side chamber 6. Is set to The rectifying passage 18 can be integrated into the first passage 8 when the shut-off position of the first on-off valve 9 is a check valve, and the suction passage 19 is also the first when the shut-off position of the second on-off valve 11 is a check valve. It can be concentrated in two passages 10.

In the actuator A configured as described above, even if the first on-off valve 9 and the second on-off valve 11 are both in the shut-off position, the rod side chamber 5, the piston side chamber 6 in the rectifying passage 18, the suction passage 19, and the discharge passage 21. And the tank 7 is made to communicate with a rosary chain. 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 or 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 not sufficient in the cylinder 2 passes from the tank 7 to the cylinder through the suction passage 19 2 is supplied. Since the variable relief valve 22 acts as a resistance against 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.

In addition, at the time of failure that prevents the actuator A from being energized, each of the first on-off valve 9 and the second on-off valve 11 takes the shut-off position, and the variable relief valve 22 has the maximum valve opening pressure. Functions as a fixed pressure control valve. Therefore, during such a failure, the actuator A automatically functions as a passive damper.

Subsequently, when causing the actuator A to exert a desired thrust in the extending direction, the controller C1 basically rotates the motor 15 to rotationally drive the pump 12 at a predetermined number of rotations, and hydraulic oil enters the cylinder 2. Supply. And let the 1st on-off valve 9 be a communicating position, and let the 2nd on-off valve 11 be a cutoff position. In this way, the rod side chamber 5 and the piston side chamber 6 are in communication with each other, and 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 generates 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 is opened 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. The actuator A then extends in the direction of extension 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 causing the actuator A to exert a thrust in a desired contraction direction, the controller C1 rotates the motor 15 and supplies the hydraulic oil from the pump 12 into the rod side chamber 5, while the first on-off valve 9 is turned on. The shut-off position is set, and the second on-off valve 11 is set to the communication position. As a result, the piston side chamber 6 and the tank 7 are brought into communication with each other and the hydraulic oil is supplied to the rod side chamber 5 from the pump 12, so that the piston 3 is pushed rightward in FIG. Demonstrate thrust. As described above, when the current of the variable relief valve 22 is adjusted, the actuator A contracts by multiplying the pressure receiving area of the piston 3 on the rod side chamber 5 side by the pressure in the rod side chamber 5 controlled by the variable relief valve 22. Demonstrate direction thrust.

Here, since the gear for pushing out the hydraulic oil in the pump 12 receives resistance according 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 from 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. As described above, the actuator A exerts a thrust according to the pressure in the rod side chamber 5 regardless of whether it extends or contracts. Therefore, if the output torque of the motor 15 is known, the thrust exerted by the actuator A Can be estimated.

Further, as can be understood from the above description, 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.

Note that the actuator A of this example not only functions as an actuator, but also functions 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 state of the motor 15. Further, when switching the actuator A from the actuator to the damper, there is no troublesome and steep switching operation of 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.

In addition, since the actuator A of this example is set to a single rod type, it is easier to secure a stroke length than the double rod type actuator, and the total length of the actuator is shortened. Mountability is improved.

In the actuator A of this example, the hydraulic oil supplied from the pump 12 and the flow of hydraulic oil by the expansion and contraction operation pass through the rod side chamber 5 and the piston side chamber 6 in order, and finally return to the tank 7. Therefore, even if gas is mixed into the rod side chamber 5 or the piston side chamber 6, the actuator A is automatically discharged to the tank 7 by the expansion / contraction operation, so that it is possible to prevent deterioration of the response of thrust generation. Therefore, when manufacturing the actuator A, it is not necessary to assemble in troublesome oil or in a vacuum environment, and advanced degassing of hydraulic oil is not required, improving productivity and reducing manufacturing cost. it can. Furthermore, even if gas is mixed in the rod side chamber 5 or the piston side chamber 6, the gas is automatically discharged to the tank 7 by the expansion / contraction operation of the actuator A. Therefore, it is necessary to frequently perform maintenance for performance recovery. The maintenance labor and cost burden can be reduced.

The configuration of the actuator A is not limited to the above. For example, a switching valve capable of communicating with the pump by selecting one of the expansion side chamber and the pressure 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 in such a configuration, the pressure of the pressure in the room to which the pump supplies hydraulic oil is received, so that the thrust of the actuator A can be estimated from the output torque of the motor driving the pump.

Subsequently, as shown in FIG. 3, the controller C1 of the present example, the target thrust Fref to be output by the actuator A based on the acceleration α in the horizontal and horizontal direction with respect to the vehicle traveling direction of the vehicle body B detected by the acceleration sensor 40. A target thrust calculation unit 41 that calculates the thrust, an estimation unit 42 that estimates the thrust of the actuator A from the current of the motor 15, and a command voltage corresponding to the current applied to the variable relief valve 22 from the deviation ε between the target thrust Fref and the estimated thrust Fm A relief valve control unit 43 for obtaining V, a motor control unit 44 for driving and controlling the motor 15 at a predetermined rotational speed, a relief valve driving unit 45 for driving the variable relief valve 22 based on the command voltage V, and a first opening / closing And an on-off valve drive unit 46 for driving and controlling 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 running included in the acceleration α detected by the acceleration sensor 40, and the target thrust that the actuator A should exert Find Fref. In this example, the target thrust calculation unit 41 is an H∞ controller, and obtains a target thrust Fref that indicates the thrust to be output by the actuator A in order to suppress the 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 number of rotations of the motor 15, feeds back the number of rotations of the motor 15 (speed feedback), and controls the motor 15 to rotationally drive the pump 12 at the predetermined number of rotations described above. To do. More specifically, the motor control unit 44 generates a current command to be given to the motor 15 from the deviation between the target rotational speed of the motor 15 and the actual rotational speed of the motor 15 for rotating the pump 12 at a predetermined rotational speed. Then, the motor 15 is controlled. Thus, 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, the estimation unit 42 first 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 pressure resistance in the rod side chamber 5, the torque of the motor 15 and the thrust of the actuator A are in a substantially proportional relationship. 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 the relationship between the torque of the motor 15 and the actual thrust of the actuator A is grasped in advance as shown in FIG. Then, if this relationship is formulated 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 using a least square method and the approximate expression may be used as an expression. In FIG. 4, the reason why the thrust of the actuator A is not exerted unless the torque of the motor 15 becomes t1 or more is due to friction of the actuator A, the pump 12 and the motor 15. Thus, 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 expansion direction or the contraction direction.

Therefore, the estimation unit 42 determines whether the direction in which the actuator A should exert the thrust is the expansion direction or the contraction direction from the sign of the target thrust Fref obtained by the target thrust calculation unit 41. 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 sign indicates the polarity that is the direction of the thrust of the actuator A, and the estimation unit 42 performs polarity determination using the sign. In this example, when the actuator A exerts a thrust in the extension direction, the target thrust Fref takes a positive value. Conversely, when the actuator A exerts a thrust in the contraction direction, the target thrust Fref takes a negative value. It is set as follows. In this way, the estimation unit 42 obtains the magnitude of the thrust of the actuator A from the current of the motor 15, performs polarity determination from the sign of the target thrust Fref, 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 the estimated thrust Fm as + a when the polarity is positive and indicates the extension direction, and the polarity Is minus and indicates the contraction direction, the estimated thrust Fm is estimated as -a. The target thrust Fref takes a negative value when the actuator A exerts a thrust in the extension direction, and the target thrust Fref takes a positive value when the actuator A exerts a thrust in the contraction direction. Also good.

The relief valve control unit 43 obtains the command voltage V from the deviation ε between the target thrust Fref obtained by the target thrust computing unit 41 and the estimated thrust Fm. In this example, the relief valve control unit 43 is a proportional compensator, and includes a deviation calculation unit 43a that calculates a deviation ε between the target thrust Fref and the estimated thrust Fm, and an absolute value processing unit 43b that performs absolute value processing on the deviation ε. And a gain multiplier 43c that multiplies the absolute value processed deviation | ε | by a proportional gain K, and a command voltage for obtaining a command voltage V to be applied to the relief valve drive unit 45 from a value | ε | K obtained by the gain multiplier 43c. And an arithmetic unit 43d. The command voltage calculation unit 43d previously stores a 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 expression. Therefore, the command voltage calculation unit 43d obtains the command voltage V to the relief valve drive unit 45 by using the above-described map or mathematical expression and using the value | ε | K as a parameter.

The relief valve drive unit 45 includes a driver circuit that drives the variable relief valve 22, and receives a command voltage V input from the relief valve control unit 43, and changes the current corresponding to the command voltage V to the variable relief valve 22. Supply to the relief valve 22. Thus, the relief valve drive unit 45 controls the valve opening pressure of the variable relief valve 22 by adjusting the current supplied to the variable relief valve 22 according to the command voltage V.

The on-off valve driver 46 determines the polarity in the expansion / contraction direction of the actuator A from the sign of the target thrust Fref obtained by the target thrust calculator 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 extending direction, the on-off valve drive unit 46 drives the first on-off valve 9 and the second on-off valve 11 to place the first on-off valve 9 in the communication position. The two on-off valve 11 is set to the cutoff 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 turn off the first on-off valve 9 in the cutoff position. The second on-off valve 11 is set to the communication position.

Note that the target thrust calculation unit 41, the estimation unit 42, and the relief valve control unit 43 in the controller C1 are not illustrated as hardware resources, but specifically, for example, A for capturing a signal output from the acceleration sensor 40. A D / D converter, a storage device such as a ROM (Read Only Memory) in which a program used for controlling the actuator A by taking the output value of the acceleration sensor 40 is stored, and based on the program And a processing 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 for the CPU. It can be realized by execution.

As described above, the railcar damping device V1 according to the first embodiment includes the actuator A that can be expanded and contracted by supplying hydraulic oil from the pump 12 driven by the motor 15, and the actuator A based on the current of the motor 15. And a controller C1 that controls the actuator A by feeding back the estimated thrust Fm estimated by the estimation unit 42.

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, and feeds back the estimated thrust Fm for closed loop control. Thus, the actuator A can be controlled.

Thus, in the railcar vibration damping device V1 of the present invention, the estimated thrust Fm is fed back without using a sensor for detecting the load of the actuator A or the pressure in the cylinder 2 when controlling the actuator A. Closed loop control can be implemented.

Therefore, according to the railway vehicle vibration damping device V1 of the present invention, the thrust of the actuator A can follow the target thrust Fref even if there is an input of a disturbance, so that a high vibration damping effect is obtained. Further, according to the railcar damping device V1 of the present invention, the system is inexpensive because it is not necessary to install a sensor for detecting the load of the actuator A or the pressure in the cylinder 2 in the closed loop control. As described above, according to the railcar damping device V1 of the present invention, a high damping effect can be obtained and the system can be inexpensive in damping the vehicle body B of the railcar. Further, according to the railcar damping device V1, tuning work is troublesome because the open-loop control requires accuracy of the valve opening pressure of the variable relief valve 22, but in the closed-loop control, the estimated thrust of the actuator A is reduced. Since the valve opening pressure of the variable relief valve 22 is automatically adjusted by feedback, the tuning operation 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 configuration of the actuator A is a configuration in which a bidirectional discharge type pump is provided in the middle of a passage communicating the extension side chamber and the pressure side chamber of the cylinder, the rotation direction of the motor is switched to extend and contract the actuator A. It comes to switch. 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. Thus, the actuator A may be configured, and the estimation unit 42 may estimate the thrust of the actuator A from only 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 Since there is an influence of inertia when switching, the thrust becomes difficult 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 only in one direction by the actuator A can be used, and the expansion / contraction direction of the actuator A can be switched. Even in this case, the thrust easily follows the target thrust Fref. As described above, according to the railcar vibration damping device V1 of this example in which the estimation unit 42 determines the polarity in the expansion / contraction direction of the actuator A based on the target thrust Fref of the actuator A, the highly responsive actuator A can be used. Thus, a higher vibration damping effect can be obtained.

In the above description, the target thrust Fref obtained by the target thrust calculation unit 41 is used for polarity determination, but the estimation unit 42 operates the first on-off valve 9 and the second on-off valve 11 as switching valves. In the situation, in this example, the polarity may be determined from the opening / closing states of the first opening / closing valve 9 and the second opening / closing valve 11. When the actuator A is extended, the on-off valve driving unit 46 supplies current to the first on-off valve 9 and does not supply current to the second on-off valve 11. Further, when the actuator A is contracted, the on-off valve driving unit 46 does not supply current to the first on-off valve 9 but supplies current to the second on-off valve 11. Thus, in this example, the polarity of the actuator A can be determined from the opening / closing states of the first opening / closing valve 9 and the second opening / closing valve 11. Therefore, when determining the polarity of the actuator A, as shown in FIG. 5, the estimation unit 42 monitors the excitation states of the first on-off valve 9 and the second on-off valve 11 that are switching valves or determines the operation status of both by the 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 a means for sensing their own position, the estimation unit 42 includes the first on-off valve 9 and the switching valve from the position obtained from the means. The operation status of the second on-off valve 11 may be grasped and the polarity may be determined. In short, the estimation unit 42 may perform polarity determination based on the operation state of the switching valve. When the railcar damping device V1 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. Also in this case, since the motor 15 that is rotationally driven only in one direction by the actuator A can be used, the highly responsive actuator A 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 performs absolute value processing of the target thrust Fref instead of the absolute value processing unit 43b that performs absolute value processing of the deviation ε. An absolute value processing unit 43f that performs absolute value processing on the thrust Fm may be provided. The deviation calculator 43a calculates a deviation ε between the absolute value processed target thrust | Fref | and the absolute value processed estimated force | Fm |, and the gain multiplier 43c and the command voltage calculator 43d thereafter What is necessary is just to process.

In addition, as described above, the relief valve control unit 43 includes only a proportional path that multiplies the proportional gain K by the deviation ε and performs only proportional compensation. And a configuration in which a differential path is added.

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

<Second Embodiment>
As shown in FIG. 7, the railcar damping device V2 of the second embodiment differs from the railcar damping device V1 of the first embodiment in the configuration of the controller C2. In the railcar vibration damping device V2 of the second embodiment, the controller C2 is a correction unit 50 that corrects the estimated thrust Fm based on the frequency of the target thrust Fref in the configuration of the controller C1 in the first embodiment. It becomes the composition which added.

As described above, the pump 12 of the actuator A in this example is a gear pump. The gear pump rotates while the two gears are engaged with each other and discharges the hydraulic oil from the discharge port while sucking the hydraulic oil from the intake port. However, the engagement resistance between the gears changes due to the backlash. On the other hand, the motor 15 that drives the pump 12 is controlled by the motor control unit 44 so as to rotate at a constant speed at the predetermined rotation speed described above. Therefore, although the pump 12 rotates at a constant speed at a predetermined rotation speed, the torque itself varies while the motor 15 rotates at a constant speed because the meshing resistance of 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 in the same manner, so that the obtained estimated thrust Fm also fluctuates so as to wave. Therefore, for example, even when the target thrust Fref takes a constant value, the deviation ε also varies because the estimated thrust Fm varies. When the value of the proportional gain K is set high, the thrust of the actuator A can easily follow the target thrust Fref. However, the command voltage V obtained by the influence of the fluctuation of the estimated thrust Fm is greatly waved. The thrust that is actually output by the actuator A in the form of a waveform also varies greatly. The fluctuation of the thrust force of the actuator A gives extra vibration to the vehicle body B, which can be a cause of deterioration of riding comfort. As described above, although there is a demand to increase the proportional gain K from the viewpoint of controllability, riding comfort is deteriorated unless the value of the proportional gain K is kept low. The deterioration of riding comfort becomes significant when the target thrust Fref is low frequency, and does not become a problem when the target thrust Fref is high frequency.

Therefore, the railway vehicle vibration damping device V2 according to 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 a 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 calculator 41 increases. The characteristic of the filter 50a is that the maximum value of the gain is 0 dB and the minimum value is −6 dB. The characteristic is set to converge to -6 dB when the frequency becomes asymptotic to a lower frequency. The gain Ky that the gain multiplication unit 50b multiplies the estimated thrust Fm after the 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 corrects the numerical value excluding the sign of the estimated thrust Fm 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 a 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 value excluding the sign of the estimated thrust Fm is corrected to be small, so that the estimated thrust Fm estimated by the estimation unit 42 fluctuates so as to wave due to 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, the numerical value excluding the sign of the estimated thrust Fm input to the relief valve control unit 43 is the actual value when the frequency of the target thrust Fref is a low frequency less than 10 Hz. Since the value is smaller than the numerical value of the thrust of the actuator A, the deviation ε obtained by the deviation calculating unit 43a becomes large. However, the value of the proportional gain K multiplied by the deviation ε by the gain multiplication unit 43c in the relief valve control unit 43 is set to a value smaller than that of the first embodiment in accordance with the increase in the deviation ε. . Even when the frequency of the target thrust Fref is 10 Hz or less and the gain in the filter 50a decreases and the deviation ε increases, the proportional gain is set so that the thrust follows the target thrust Fref without causing the command voltage V to be excessive. Both K and gain Ky are tuned.

As described above, 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 the correction can be suppressed. 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, the command voltage V becomes excessive when the correction unit 50 corrects the numerical value excluding the sign of the estimated thrust Fm to increase the deviation ε. However, when the frequency of the target thrust Fref is a high frequency of 10 Hz or higher, 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 no significant 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 when the frequency of the target thrust Fref is 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 follow the target thrust Fref with high accuracy.

As described above, the railcar 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, and thus uses a gear pump for the pump 12. Even so, the driving force of the actuator A can be made to follow the target thrust Fref in the entire frequency range of the target thrust Fref, so that the riding comfort in the vehicle can be further improved.

Therefore, according to the railcar damping device V2 in the second embodiment, not only the operational effect of the railcar damping device V1 in the first embodiment is exhibited, 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 processing that causes a time delay, so that it is not necessary to cause deterioration in control performance.

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

Further, the railcar vibration damping device V2 in the second embodiment includes a proportional path that multiplies the proportional gain K by the deviation ε, and is therefore proportional to the gain Ky when the correction unit 50 corrects the estimated thrust Fm. Tuning of the two gains K can be performed, and the riding comfort of the vehicle can be further effectively improved by causing the thrust of the actuator A to follow the target thrust Fref with high accuracy in the entire frequency range of the target thrust Fref.

Note that the correction unit 50 may be a low-pass filter, and the estimated thrust Fm estimated due to the influence of the structure of the pump 12 may fluctuate so as to wave. In this way, since the vibration component of the estimated thrust Fm can be removed, the proportional gain K does not need to be set small, and the thrust of the actuator A can be controlled so as not to be vibrational, thereby improving riding comfort.

The preferred embodiments of the present invention have been described in detail above, but modifications, changes, and changes can be made without departing from the scope of the claims.

This application claims priority based on Japanese Patent Application No. 2017-040420 filed with the Japan Patent Office on March 3, 2017, the entire contents of which are incorporated herein by reference.

Claims

A railway vehicle damping device,
An actuator that can be expanded and contracted by supplying a working liquid from a pump driven by a motor;
And a controller for controlling the actuator by feeding back the estimated thrust estimated by the estimation unit, and a controller for controlling the actuator by feeding back the estimated thrust estimated by the estimation unit.
A vibration damping device for a railway vehicle according to claim 1,
An acceleration sensor,
The controller obtains a target thrust of the actuator based on the acceleration detected by the acceleration sensor,
The said estimation part determines the polarity which is the expansion-contraction direction of the said actuator based on the said target thrust.
A vibration damping device for a railway vehicle according to claim 1,
The actuator has a switching valve for switching the expansion and contraction direction,
The said estimation part determines the polarity which is the expansion-contraction direction of the said actuator based on the operation condition of the said switching valve.
A vibration damping device for a railway vehicle according to claim 2,
A railcar vibration control device comprising a correction unit that corrects the estimated thrust based on the frequency of the target thrust.
A vibration damping device for a railway vehicle according to claim 3,
An acceleration sensor,
The controller obtains a target thrust of the actuator based on the acceleration detected by the acceleration sensor,
A railcar vibration control device comprising a correction unit that corrects the estimated thrust based on the frequency of the target thrust.
The vibration damping device for a railway vehicle according to claim 5,
The controller has a proportional path to which a deviation between the target thrust and the estimated thrust is input,
The railroad vehicle vibration damping device that corrects the estimated thrust by processing the estimated thrust with a filter having a characteristic that gain increases as the frequency of the target thrust increases.
The vibration damping device for a railway vehicle according to claim 5,
The said correction | amendment part is a vibration suppression device for rail vehicles which carries out the low-pass filter process of the said estimated thrust.
A vibration damping device for a railway vehicle according to claim 1,
The actuator is
A cylinder,
A piston 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;
A tank,
The pump capable of sucking the working liquid from the tank and supplying the working liquid to the rod side chamber;
The motor for driving the pump;
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 a 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 toward the rod side chamber;
A railcar damping device comprising: a suction passage that allows only a flow of hydraulic oil from the tank toward the piston side chamber.