Patent Application
20190111950
The present invention relates to modification of a railway vehicle vibration damping device.
For example, a conventionally known railway vehicle vibration damping device is interposed between a vehicle body and a truck of a railway vehicle to reduce vibration in the right and left directions relative to the traveling direction of the vehicle body.
More specifically, for example, as disclosed in JP 2010-65797 A, such a railway vehicle vibration damping device includes: an actuator that includes a cylinder, a piston slidably inserted into the cylinder and partitioning inside of the cylinder into a rod side chamber and a piston side chamber, and a rod inserted into the cylinder and coupled with the piston, and that is interposed between the vehicle body and the truck; a tank; a first on-off valve provided halfway through a first path through which the rod side chamber and the piston side chamber are communicated with each other; a second on-off valve provided halfway through a second path through which the piston side chamber and the tank are communicated with each other; a pump configured to supply hydraulic oil to the rod side chamber; a motor configured to drive the pump; a discharge path connecting the rod side chamber to the tank; and a variable relief valve that is provided halfway through the discharge path and the valve opening pressure of which is changeable. Thrust can be exerted in expansion and contraction by driving the pump, the first on-off valve, the second on-off valve, and the variable relief valve, thereby reducing vibration of the vehicle body.
In the conventional railway vehicle vibration damping device, the motor drives the pump at a constant rotational speed (rotations per unit time), the first on-off valve, the second on-off valve, and the variable relief valve are driven as appropriate in accordance with the vibration status of the vehicle body, hydraulic pressure is used to obtain thrust for reducing vibration of the vehicle body, thereby damping vibration of the vehicle body.
In such a conventional railway vehicle vibration damping device, the speed of expansion and contraction of the actuator increases when the vehicle body vibrates at high speed, for example, when the railway vehicle passes through a point.
The pressure in the actuator temporarily becomes excessive in some cases when the actuator expands or contracts at high speed. To rotate the pump at a constant speed in such a case, the motor needs to generate torque exceeding the rated torque thereof.
Typically, at activation or the like, the motor needs to output torque (hereinafter referred to as “excessive torque”) exceeding the rated torque, and thus is capable of outputting the excessive torque. However, when the motor continuously outputs the excessive torque, the motor is seized, which is called burnout. Thus, a time in which the excessive torque is output is restricted to avoid burnout when such excessive torque is output, thereby protecting the motor from an overload. Specifically, a motor driver configured to drive the motor monitors a time in which the motor outputs the excessive torque, and executes emergency stop processing of the motor when the time in which the excessive torque is output reaches a restriction time.
With this configuration, at point passing or the like, the conventional railway vehicle vibration damping device performs emergency stop of the motor to protect the motor and interrupts the vibration damping control in some cases. Reset needs to be performed to activate the motor again, and thus the vibration damping control cannot be restarted while the railway vehicle travels.
An object of the present invention is to provide a railway vehicle vibration damping device capable of protecting a motor from an overload and avoiding interruption of vibration damping control.
A railway vehicle vibration damping device according to the present invention includes: an actuator including a cylinder body configured to expand and contract when supplied with hydraulic fluid, a pump configured to supply the hydraulic fluid to the cylinder body, and a motor configured to drive the pump, and installed on a railway vehicle; and a control unit configured to control the motor. The control unit stops the motor when a current value to be provided to the motor becomes equal to or larger than a current threshold at which a restriction torque set to a value exceeding a rated torque of the motor is output.
The present invention will be described below based on an embodiment illustrated in the accompanying drawings. In the present example, a railway vehicle vibration damping device 1 according to the embodiment is used as a vibration damping device for a vehicle body B of a railway vehicle, and includes an actuator A installed between a truck T and the vehicle body B, and a control unit C as illustrated in
In the present example, as illustrated in
In the present example, the rod side chamber 5 and the piston side chamber 6 are filled with the hydraulic oil as hydraulic fluid, and the tank 7 is filled with gas in addition to the hydraulic oil. The tank 7 does not need to be filled with compressed gas to achieve a pressurized state. The hydraulic fluid may be any other liquid in place of the hydraulic oil.
The fluid pressure circuit HP includes a first on-off valve 9 provided halfway through a first path 8 through which the rod side chamber 5 and the piston side chamber 6 are communicated with each other, and a second on-off valve 11 provided halfway through a second path 10 through which the piston side chamber 6 and the tank 7 are communicated with each other.
Basically, the cylinder body Cy expands when the pump 12 is driven while the first path 8 is made into a communicating state by the first on-off valve 9 and the second on-off valve 11 is closed, or the cylinder body Cy contracts when the pump 12 is driven while the second path 10 is made into a communicating state by the second on-off valve 11 and the first on-off valve 9 is closed.
Each component of the actuator A will be described below in detail. The cylinder 2 has a tubular shape, the right end of which in
The gap between the outer periphery of the rod guide 14 and the cylinder 2 is sealed by a seal member (not illustrated) to maintain the cylinder 2 in a sealed state. The rod side chamber 5 and the piston side chamber 6, which are partitioned by the piston 3 in the cylinder 2, are filled with the hydraulic oil as described above.
In the cylinder body Cy, the cross-sectional area of the rod 4 is set to one-half of the cross-sectional area of the piston 3 so that the pressure receiving area of the piston 3 on the rod side chamber 5 side is half of the pressure receiving area thereof on the piston side chamber 6 side. With this configuration, when the pressure in the rod side chamber 5 is set to be the same between expansion and contraction operations, the same magnitude of thrust is generated at both of expansion and contraction, and accordingly, the amount of hydraulic oil per displacement amount of the cylinder body Cy is the same between expansion and contraction.
Specifically, at the expansion operation of the cylinder body Cy, the rod side chamber 5 and the piston side chamber 6 are communicated with each other. Accordingly, the pressure in the rod side chamber 5 and the pressure in the piston side chamber 6 become equal to each other, and the actuator A generates thrust in a magnitude equal to a value obtained by multiplying, by the pressure, the difference in the pressure receiving area of the piston 3 between the rod side chamber 5 side and the piston side chamber 6 side. At the contraction operation of the cylinder body Cy, the communication between the rod side chamber 5 and the piston side chamber 6 is cut off so that the piston side chamber 6 is communicated with the tank 7 Accordingly, the actuator A generates thrust in a magnitude equal to a value obtained by multiplying, by the pressure in the rod side chamber 5, the pressure receiving area of the piston 3 on the rod side chamber 5 side.
In other words, the thrust generated by the actuator A at expansion and contraction has a value obtained by multiplying one-half of the cross-sectional area of the piston 3 with the pressure in the rod side chamber 5. Thus, the thrust of the actuator A can be controlled by controlling the pressure in the rod side chamber 5 at the expansion and contraction operations. In addition, since, in the actuator A in the present example, the pressure receiving area of the piston 3 on the rod side chamber 5 side is set to one-half of the pressure receiving area thereof on the piston side chamber 6 side, the pressure in the rod side chamber 5 is the same between expansion and contraction when the same thrust is generated at expansion and contraction, which simplifies control. In addition, since the amount of hydraulic oil per displacement amount is the same, the same response characteristic is obtained between expansion and contraction. When the pressure receiving area of the piston 3 on the rod side chamber 5 side is not set to one-half of the pressure receiving area thereof on the piston side chamber 6 side, the thrust of the actuator A at expansion and contraction can be still controlled by controlling the pressure in the rod side chamber 5.
The left end of the rod 4 in
The rod side chamber 5 and the piston side chamber 6 are communicated with each other through the first path 8, and the first on-off valve 9 is provided halfway through the first path 8. The first path 8, through which the rod side chamber 5 and the piston side chamber 6 are communicated with each other outside of the cylinder 2, may be provided to the piston 3.
The first on-off valve 9 is an electromagnetic on-off valve, and has a communicating position at which the first path 8 is opened so that the rod side chamber 5 and the piston side chamber 6 are communicated with each other, and a cutoff position at which the first path 8 is cut off so that the communication between the rod side chamber 5 and the piston side chamber 6 is cut off. The first on-off valve 9 is set to the communicating position when energized, or to the cutoff position when not energized.
The piston side chamber 6 and the tank 7 are communicated with each other through the second path 10, and the second on-off valve 11 is provided halfway through the second path 10. The second on-off valve 11 is an electromagnetic on-off valve, and has a communicating position at which the second path 10 is opened so that the piston side chamber 6 and the tank 7 are communicated with each other, and a cutoff position at which the second path 10 is cut off so that the communication between the piston side chamber 6 and the tank 7 is cut off. The second on-off valve 11 is set to the communicating position when energized, or to the cutoff position when not energized.
The pump 12 is a pump driven by the motor 15 to discharge the hydraulic oil only in one direction. The pump 12 has a discharge port communicated with the rod side chamber 5 through a supply path 16, and has a suction port communicated with the tank 7. When driven by the motor 15, the pump 12 sucks the hydraulic oil from the tank 7 and supplies the hydraulic oil to the rod side chamber 5.
Since the pump 12 discharges the hydraulic oil only in one direction as described above and thus does not switch the rotational direction thereof, the pump 12 has no problem such as discharge amount change at rotation switching, and thus may be, for example, an inexpensive gear pump. In addition, since the rotational direction of the pump 12 is constantly in the same one direction, the motor 15 as a drive source configured to drive the pump 12 is not requested to have a high response characteristic to the rotation switching, and thus may be an inexpensive motor. A check valve 17 is provided halfway through the supply path 16 to prevent backflow of the hydraulic oil from the rod side chamber 5 to the pump 12.
The fluid pressure circuit HC in the present example includes, in addition to the above-described configuration, a discharge path 21 connecting the rod side chamber 5 and the tank 7, and a variable relief valve 22 that is provided halfway through the discharge path 21 and valve opening pressure of which is changeable.
In the present example, the variable relief valve 22 is a proportional electromagnetic relief valve, and the valve opening pressure thereof is adjustable in accordance with the amount of supplied current: the valve opening pressure is minimized when the maximum amount of current is supplied, or the valve opening pressure is maximized when no current is supplied.
When the discharge path 21 and the variable relief valve 22 are provided in this manner, at expansion and contraction operations of the cylinder body Cy, 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 controlled by the amount of current supplied to the variable relief valve 22. When the discharge path 21 and the variable relief valve 22 are provided, sensors and the like necessary for adjusting the thrust of the actuator A are not needed, and the motor 15 does not need to be highly precisely controlled to adjust the discharge flow rate of the pump 12. This leads to cost reduction of the railway vehicle vibration damping device 1, and leads to a system that is robust as hardware and software.
When the first on-off valve 9 is set to the communicating position and the second on-off valve 11 is set to the cutoff position or when the first on-off valve 9 is set to the cutoff position and the second on-off valve 11 is set to the communicating position, the actuator A exerts damping force only for any one of expansion and contraction irrespective of the drive status of the pump 12. Thus, for example, when the direction of the damping force is in a direction in which the vehicle body B is vibrated due to vibration of the truck T of the railway vehicle, the actuator A can be a single-side acting damper so that the damping force is not applied in such a direction. Accordingly, the actuator A can easily achieve Karnopp control, and thus can function as a skyhook semi-active damper.
When a proportional electromagnetic relief valve, the valve opening pressure of which is proportionally changed by the amount of supplied current is used as the variable relief valve 22, the valve opening pressure can be easily controlled. However, the variable relief valve is not limited to a proportional electromagnetic relief valve, but may be any variable relief valve, the valve opening pressure of which is adjustable is applicable.
The variable relief valve 22 opens the discharge path 21 when excessive input is made to the cylinder body Cy in the directions of expansion and contraction so that the pressure in the rod side chamber 5 exceeds the valve opening pressure, irrespective of whether the first on-off valve 9 are the second on-off valve 11 are opened or closed. In this manner, when the pressure in the rod side chamber 5 becomes equal to or larger 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 excessive pressure in the cylinder 2, thereby protecting the entire system of the actuator A. Thus, the system can be protected by providing the discharge path 21 and the variable relief valve 22.
The fluid pressure circuit HC in the actuator A in the present example includes a rectification path 18 that only allows flow of the hydraulic oil from the piston side chamber 6 to the rod side chamber 5, and a suction path 19 that only allows flow of the hydraulic oil from the tank 7 to the piston side chamber 6. With this configuration, in the actuator A in the present example, the hydraulic oil is pushed out of the cylinder 2 when the cylinder body Cy expands or contracts while the first on-off valve 9 and the second on-off valve 11 are closed. The variable relief valve 22 applies resistance against flow of the hydraulic oil discharged out of the cylinder 2, and thus the actuator A in the present example functions as a uniflow damper when the first on-off valve 9 and the second on-off valve 11 are closed.
More specifically, the rectification path 18 is set to be a one-way path through which the piston side chamber 6 and the rod side chamber 5 are communicated with each other, halfway through which a check valve 18a is provided, and that only allows flow of the hydraulic oil from the piston side chamber 6 to the rod side chamber 5. The suction path 19 is set to be a one-way path through which the tank 7 and the piston side chamber 6 are communicated with each other, halfway through which a check valve 19a is provided, and that only allows flow of the hydraulic oil from the tank 7 to the piston side chamber 6. The rectification path 18 can be collected to the first path 8 when the cutoff position of the first on-off valve 9 is set to a check valve, and the suction path 19 can be collected to the second path 10 when the cutoff position of the second on-off valve 11 is set to a check valve.
In the actuator A thus configured, when the first on-off valve 9 and the second on-off valve 11 are each set to the cutoff position, the rod side chamber 5, the piston side chamber 6, and the tank 7 are sequentially communicated with each other through the rectification path 18, the suction path 19, and the discharge path 21. The rectification path 18, the suction path 19, and the discharge path 21 are set to be one-way. With this configuration, when the cylinder body Cy is expanded or contracted by external force, the hydraulic oil is always discharged from the cylinder 2 and returned to the tank 7 through the discharge path 21, and hydraulic oil in an amount compensating the hydraulic oil discharged from the cylinder 2 is supplied from the tank 7 into the cylinder 2 through the suction path 19. The variable relief valve 22 applies resistance against this flow of the hydraulic oil so that the pressure in the cylinder 2 is adjusted to the valve opening pressure, and thus the actuator A functions as a passive uniflow damper.
At failure of energization to each instrument of the actuator A, the first on-off valve 9 and the second on-off valve 11 are each set to the cutoff position, and the variable relief valve 22 functions as a pressure control valve, the valve opening pressure of which is fixed to maximum. Accordingly, at such failure, the actuator A automatically functions as a passive damper.
When causing the actuator A to exert desired thrust in the contraction direction, the control unit C basically sets the first on-off valve 9 to the communicating position and sets the second on-off valve 11 to the cutoff position while rotating the motor 15 to supply the hydraulic oil from the pump 12 into the cylinder 2. In this manner, the rod side chamber 5 and the piston side chamber 6 are communicated with each other and supplied with the hydraulic oil from the pump 12, and the piston 3 is pushed to the left in
When causing the actuator A to exert desired thrust in the contraction direction, the control unit C sets the first on-off valve 9 to the cutoff position and sets the second on-off valve 11 to the communicating position while rotating the motor 15 to supply the hydraulic oil from the pump 12 into the rod side chamber 5. In this manner, the piston side chamber 6 and the tank 7 are communicated with each other and the hydraulic oil is supplied from the pump 12 to the rod side chamber 5, and the piston 3 is pushed to the right in
The actuator A not only functions as an actuator but also functions as a damper only through opening and closing of the first on-off valve 9 and the second on-off valve 11 irrespective of the drive status of the motor 15. No cumbersome and abrupt switching operation of the first on-off valve 9 and the second on-off valve 11 needs to be performed to switch the actuator A from an actuator to a damper, and thus a highly responsive and reliable system can be achieved.
Since the actuator A in the present example is set to be a single-rod type, it is easier to have a sufficient stroke length for the actuator A than for a both-rod type actuator, which leads to a shorter total length of the actuator A and facilitates its mounting onto the railway vehicle.
The flow of the hydraulic oil due to hydraulic oil supply from the pump 12 and the expansion and contraction operations in the actuator A in the present example sequentially passes through the rod side chamber 5 and the piston side chamber 6 and finally flows back to the tank 7 Thus, any gas mixed into the rod side chamber 5 or the piston side chamber 6 is autonomously discharged to the tank 7 through the expansion and contraction operations of the cylinder body Cy, thereby preventing degradation of the response characteristic of thrust generation. This eliminates the need to perform cumbersome assembly in oil or under vacuum environment at manufacturing of the actuator A, and highly precise reaeration of the hydraulic oil, which leads to improvement of productivity and reduction of manufacturing cost. In addition, since any gas mixed into the rod side chamber 5 or the piston side chamber 6 is autonomously discharged to the tank 7 through the expansion and contraction operations of the cylinder body Cy, maintenance for performance recovery does not need to be frequently performed, which leads to reduction of loads on maintenance work and cost.
The control unit C will be described next. As illustrated in
As illustrated in
In the present example, the control force calculation unit 421 is an H[infinity] controller, and calculates, from the lateral acceleration a, the control force F instructing thrust to be output from the actuator A to reduce vibration of the vehicle body B. The control force F is provided with a positive or negative sign depending on the direction thereof, and the sign indicates the direction of thrust to be output from the actuator A. When having received inputting of the control force F, the on-off valve drive unit 425 drives opening and closing of the first on-off valve 9 and the second on-off valve 11 by supplying current thereto or stops current supply thereto in accordance with the sign of the control force F. More specifically, when the contraction direction of the actuator A is defined to be positive and the contraction direction thereof is defined to be negative, the on-off valve drive unit 425 operates as follows. When the sign of the control force F is positive, thrust is to be exerted by the actuator A in the contraction direction, and thus the on-off valve drive unit 425 sets the first on-off valve 9 to the communicating position and sets the second on-off valve 11 to the cutoff position. Accordingly, the hydraulic oil is supplied from the pump 12 to both of the rod side chamber 5 and the piston side chamber 6, and thus the actuator A exerts thrust in the contraction direction. When the sign of the control force F is negative, thrust is exerted by the actuator A in the contraction direction, and thus the on-off valve drive unit 425 sets the first on-off valve 9 to the cutoff position and sets the second on-off valve 11 to the communicating position. Accordingly, the hydraulic oil is supplied from the pump 12 only to the rod side chamber 5 so that the piston side chamber 6 and the tank 7 are communicated with each other, and thus the actuator A exerts thrust in the contraction direction.
In the present example, the control force calculation unit 421 calculates the control force F only from the lateral acceleration a. However, control force for reducing sway of the vehicle body B and control force for reducing yaw thereof may be separately calculated based on the sway acceleration and yaw acceleration of the vehicle body B, and added to obtain the control force F.
The motor current value calculation unit 422 receives inputting of a rotational speed and a current amount from a sensor 43 configured to sense the rotational speed of the motor 15 and a sensor 44 configured to sense the amount of current flowing to the motor 15, respectively, and monitors the rotational speed of the motor 15 and the amount of current flowing to the motor 15. The motor current value calculation unit 422 includes a speed loop and a current loop, and feeds back the rotational speed of the motor 15 and the current amount thereto to calculate the current value I to be provided to the motor 15 to drive the motor 15 at the predetermined rotational speed. When the motor 15 is a brushless motor, the rotational speed of the motor 15 may be sensed normally by using, as the sensor 43, for example, a resolver or a Hall sensor for sensing the electric angle of the motor 15. The amount of current to the motor 15 may be sensed by using, as the sensor 44, a current sensor normally included in the motor 15. The predetermined rotational speed may be determined in advance to be a value optimum for vibration damping of a railway vehicle to which the railway vehicle vibration damping device 1 is applied. Specifically, the motor current value calculation unit 422 sets the predetermined rotational speed as a target rotational speed, calculates a target current value based on deviation of the rotational speed of the motor 15 from the target rotational speed, and calculates the current value I to be provided to the motor 15 based on deviation of the target current value from the amount of current actually flowing to the motor 15.
The motor current value determination unit 423 compares the current value I calculated by the motor current value calculation unit 422 and a current threshold Iα to be described later. As a result of the comparison, the motor current value determination unit 423 sets the current value I to the current value Ie, which is finally provided to the motor 15, when the current value I is smaller than the current threshold Iα, or sets zero to the current value Ie, which is finally provided to the motor 15, when the current value I is equal to or larger than the current threshold Iα. The current value Ie is command current provided to the motor 15. The current threshold Iα is set to a current value at which the motor 15 outputs a restriction torque exceeding a rated torque. The restriction torque may be set in accordance with specifications of the railway vehicle vibration damping device 1, but is set to be, for example, in the range of 200% to 500% of the rated torque. When torque needed to be output at activation of the motor 15 exceeds the above-described restriction torque, the current value I calculated by the motor current value calculation unit 422 may be set to the final current value Ie without performing the processing at the motor current value determination unit 423 at activation of the motor 15. When the current value I becomes smaller than the current threshold Iα after the current value I becomes equal to or larger than the current threshold Iα and the current value Ie is set to zero, the motor current value determination unit 423 sets the current value I to the current value Ie.
The processing at the motor current value calculation unit 422 and the motor current value determination unit 423 will be described next with reference to a flowchart illustrated in
Subsequently, the relief valve current value calculation unit 424 calculates the relief valve current value IR to be supplied to the variable relief valve 22 based on the control force F calculated as described above. The valve opening pressure of the variable relief valve 22 changes proportionally to the supplied current amount, and the variable relief valve 22 has a characteristic of pressure override that a pressure loss increases in accordance with a passing flow rate. Since the rotational speed of the motor 15 constantly rotates at the predetermined rotational speed, the amount of hydraulic oil passing through the variable relief valve 22 can be approximately assumed, and thus the relief valve current value calculation unit 424 calculates the relief valve current value IR with the pressure override taken into account.
In the present example, the relief valve control unit 426 is a driver configured to drive a solenoid (not illustrated) of the variable relief valve 22, and supplies the amount of current at the relief valve current value IR to the variable relief valve 22 when having received inputting of the relief valve current value IR.
The motor driver 427 supplies current to the motor 15. When the motor 15 is driven, the pump 12 rotates. In the present example, when having received inputting of the current value Ie, the motor driver 427 performs PWM control of the motor 15 to drive the motor 15 so that the amount of current flowing to the motor 15 is equal to a current amount instructed by the current value Ie.
Although not illustrated, the control unit C may specifically include, as hardware resources, for example, an A/D converter for acquiring signals output from the acceleration sensor 40 and the sensors 43 and 44; a storage device such as a read only memory (ROM) configured to acquire the lateral acceleration a filter through the band-pass filter 41 and store a computer program used in processing necessary for controlling the actuator A; an arithmetic device such as a central processing unit (CPU) configured to execute processing based on the computer program; and a storage device such as a random access memory (RAM) configured to provide a storage region to the CPU. Each component of the control processing unit 42 of the control unit C may be achieved by the CPU executing the computer program. The band-pass filter 41 may be achieved by the CPU executing the computer program.
In this manner, the railway vehicle vibration damping device 1 sets zero to the current value Ie provided to the motor 15 when the current value I to be provided to the motor 15 becomes equal to or larger than the current threshold Iα, and stops the motor 15. Since current at the current threshold Iα causes the motor 15 to output torque exceeding the rated torque, the current applies an overload on the motor 15 and burns out the motor 15 when continuously provided to the motor 15 for a long time. However, the railway vehicle vibration damping device 1 according to the present invention stops the motor 15 when the current value I to be provided to the motor 15 becomes equal to or larger than the current threshold Iα, thereby protecting the motor 15. Since the railway vehicle vibration damping device 1 according to the present invention stops the motor 15 when the current value I becomes equal to or larger than the current threshold Iα, the motor 15 can be stopped before emergency stop control incorporated in the motor driver 427 in advance to protect the motor 15 from an overload is executed. Thus, in the railway vehicle vibration damping device 1 according to the present invention, current flowing to the motor 15 can be restricted before emergency stop of the motor 15, and thus the motor 15 is constantly maintained in a drivable state to avoid incapability of executing vibration damping control while the railway vehicle is traveling. In this manner, according to the railway vehicle vibration damping device 1 according to the present invention, the motor 15 can be protected from an overload, and interruption of the vibration damping control can be avoided.
The motor current value determination unit 423 sets the current value I to the current value Ie when the current value I is smaller than the current threshold Iα after the current value I becomes equal to or larger than the current threshold Iα and the current value Ie is set to zero. In this manner, in the railway vehicle vibration damping device 1 in the present example, drive of the motor 15 is resumed when the current value I becomes smaller than the current threshold Iα, and thus the vibration damping control is immediately resumed when stop of the motor 15 becomes unnecessary, thereby continuously achieving a vibration damping effect.
The railway vehicle vibration damping device 1 in the present example includes: the cylinder body Cy including the cylinder 2, the piston 3, and the rod 4; the tank 7; the pump 12 configured to supply the hydraulic oil to the rod side chamber 5; the motor 15 configured to drive the pump 12; the first on-off valve 9 provided halfway through the first path 8 through which the rod side chamber 5 and the piston side chamber 6 are communicated with each other; the second on-off valve 11 provided halfway through the second path 10 through which the piston side chamber 6 and the tank 7 are communicated with each other; the discharge path 21 connecting the rod side chamber 5 and the tank 7; the variable relief valve 22 that is provided halfway through the discharge path 21 and the valve opening pressure of which is changeable; the rectification path 18 that only allows flow of the hydraulic oil from the piston side chamber 6 to the rod side chamber 5; and the suction path 19 that only allows flow of the hydraulic oil from the tank 7 to the piston side chamber 6. In the railway vehicle vibration damping device 1 thus configured, since the actuator A functions as a skyhook semi-active damper while the motor 15 is stopped, the vibration damping effect is maintained during the stop of the motor 15.
As described above, the railway vehicle vibration damping device 1 stops the motor 15 when the current value I becomes equal to or larger than the current threshold Iα. However, a time in which the current value I is equal to or larger than the current threshold Iα may be counted to stop the motor 15 before the emergency stop control for protecting the motor 15 from an overload is executed. Specifically, the railway vehicle vibration damping device 1 counts the time in which the current value I is equal to or larger than the current threshold Iα, and stops the motor 15 when the counted time reaches a predetermined time. The current value of the motor 15 and time are monitored, and the emergency stop control for protecting the motor 15 from an overload is executed when current flows to the motor 15 for a time long enough to cause burnout. The predetermined time is set to be shorter than a time until the emergency stop control is executed since the current value of the motor 15 has exceeded the current threshold Iα, which potentially causes burnout. This allows the railway vehicle vibration damping device 1 to stop the motor 15 before the emergency stop control for protecting the motor 15 from an overload is executed. A time until the motor 15 burns out is different depending on the magnitude of the current value, and thus the predetermined time may be changed in accordance with the magnitude of the current value. Since the time in which the current value I is equal to or larger than the current threshold Iα is counted, the motor 15 is not stopped when the current value temporarily becomes equal to or larger than the current threshold Iα.
While the preferred embodiment of the present invention have been described in detail, modifications, variations, and changes can be made without departing from the scope of the claims.
The present application claims priority based on Japanese Patent Application No. 2016-158559 filed on Aug. 12, 2016, the Japan Patent Office, the entire contents of this application being incorporated herein by reference.
| Number | Date | Country | Kind |
|---|---|---|---|
| 2016-158559 | Aug 2016 | JP | national |
| Filing Document | Filing Date | Country | Kind |
|---|---|---|---|
| PCT/JP2017/015734 | 4/19/2017 | WO | 00 |
