JP2018122626A - Stationary acceleration speed detection device and vibration damping device for railway vehicle - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a stationary acceleration speed detection device which can detect the stationary acceleration speed with high accuracy, and a vibration damping device for railway vehicle which can improve the riding comfort at the travel in the curve section.SOLUTION: A stationary acceleration speed detection device D obtains the stationary acceleration speed αc on the basis of the filtered yaw acceleration speed ω' and filtered sway acceleration speed β' subjected to the low-pass filter processing. A vibration damping device V for railway vehicle obtains the control forces Ff, Fr on the basis of the centering force fn and the suppression forces ff, fr obtained on the basis of the stationary acceleration speed αc when the stationary acceleration speed αc detected by the stationary acceleration speed detection device D is equal to or greater than a centering threshold α1.SELECTED DRAWING: Figure 3

Description

The present invention relates to improvements in a steady acceleration detection device and a railcar vibration damping device.

  The railway vehicle is provided with a double-acting actuator interposed between the vehicle body and the carriage and a controller for controlling the actuator, and the railway vehicle control system suppresses vibration in the lateral direction with respect to the traveling direction of the vehicle body. A vibration device may be provided.

  Such a railcar damping device detects the sway acceleration and yaw acceleration of the vehicle body of the railcar, controls the actuator by acceleration feedback, and suppresses left-right movement of the vehicle body (see, for example, Patent Document 1).

  In addition, the railcar vibration damping device detects the steady acceleration by low-pass filtering the sway acceleration to determine whether the railway vehicle is traveling in a curved section, and is suitable for straight section traveling. The vibration control effect is enhanced by switching between vibration control and vibration control suitable for running in a curved section. The steady acceleration refers to centrifugal acceleration that constantly acts on the vehicle body when the railway vehicle travels in a curved section.

JP 2014-46714 A

  However, in the above-described railcar vibration damping device, since the sway acceleration is subjected to low-pass filter processing, the phase of the obtained steady acceleration is delayed from the actual steady acceleration, so that the steady acceleration cannot be obtained with high accuracy.

  Therefore, there is a tendency that it is determined that the vehicle is traveling in a straight section at the entrance of the curved section and is still determined as a curved section at the exit of the curved section, and the curved section cannot be accurately determined.

  Therefore, an object of the present invention is to provide a steady acceleration detecting device capable of detecting steady acceleration with high accuracy and a railcar vibration damping device capable of improving riding comfort during traveling in a curved section.

  The steady acceleration detection device of the present invention obtains a steady acceleration based on a filtered sway acceleration obtained by processing a sway acceleration with a low pass filter for sway acceleration and a post-filtered yaw acceleration obtained by processing a yaw acceleration by a low pass filter for yaw acceleration. Yes. At the entrance and exit of the curve section, the yaw acceleration acts on the vehicle body at a timing earlier than the low-frequency component of the sway acceleration obtained by filtering with the sway acceleration low-pass filter, so it is obtained by filtering with the sway acceleration low-pass filter. If the yaw acceleration ω is added in addition to the sway acceleration, the phase delay is compensated, and a steady acceleration with little or no phase delay with respect to the actual steady acceleration can be obtained.

  Further, when the steady acceleration detecting device obtains the steady acceleration by subtracting the yaw acceleration from the sway acceleration, the phase delay of the obtained steady acceleration with respect to the actual steady acceleration can be eliminated using a simple calculation.

  Furthermore, when the steady acceleration detected by the steady acceleration detection device is equal to or greater than the centering threshold, the railcar damping device obtains the control force based on the centering force and the suppression force obtained based on the steady acceleration. In the railcar vibration damping device configured as described above, the necessity of exhibiting the centering force is determined based on the value of steady acceleration, and a displacement sensor is not required. And, according to this railcar damping device, it is possible to exert the restraining force and centering force to suppress vibration during traveling in a curved section, and it is possible to suppress the vehicle body from contacting the stopper and compressing it to the maximum. In this case, it is possible to suppress transmission of vibration from the cart side to the vehicle body.

  Further, the railcar vibration damping device may obtain the centering force with the upper limit of the centering force as the maximum value of the force that the actuator can exert when the motor drives the pump with the rated torque. In the railcar damping device configured as described above, even if the actuator outputs only the centering force, the remaining force remains until the maximum torque that the motor can output. It is also possible to output a suppression force for suppressing vibrations. Therefore, according to the railcar vibration damping device, it is possible to exert a restraining force that suppresses vibrations of the vehicle body while exhibiting a centering force to return the vehicle body to the neutral position when traveling in a curved section, and to improve riding comfort during traveling in a curved section. It can be further improved.

  Furthermore, the railway vehicle vibration damping device has a straight section control unit and a curved section control unit to obtain the suppression force, and the curve determination is made from the absolute value of the steady acceleration that is larger than the centering threshold and less than the curve determination threshold. When the threshold value is greater than or equal to the threshold value, the suppression force obtained by the straight line section control unit is switched to the suppression force obtained by the curve section control unit. May be switched to the suppression force required by the straight section controller. According to the vibration damping device for a railway vehicle configured as described above, an optimal suppression force can be exhibited at the entrance and exit of the curved section, and a high vibration suppression effect can be obtained regardless of the travel section, thereby improving riding comfort.

  Further, the railcar damping device fades out and switches the suppression force selected before switching when switching between the suppression force obtained by the straight section control unit and the suppression force obtained by the curved section control unit. The suppression force to be selected later may be faded in. According to the railcar damping device configured as described above, the stability of the control is improved because the value of the suppression force does not change suddenly when switching the suppression force for the straight section and the suppression force for the curved section. To do.

  According to the steady acceleration detecting apparatus of the present invention, the steady acceleration can be detected with high accuracy. Moreover, according to the railcar damping device of the present invention, it is possible to improve riding comfort during traveling in a curved section.

It is a top view of a railway vehicle carrying a railcar damping device. It is detail drawing of an actuator. It is a control block diagram of a controller in a railcar vibration damping device. It is a control block diagram of the control calculation part of the controller in a railcar damping device. It is a control block diagram of the yaw suppression force calculating part in a control calculating part. It is the figure which showed the gain for straight areas, and the gain for curve areas. It is a control block diagram of the sway suppression force calculating part in a control calculating part. It is a control block diagram of the centering force calculation unit in the control calculation unit. It is a control block diagram of the control force calculation part in a control calculation part. It is the flowchart which showed the process sequence in a control calculating part.

  The present invention will be described below based on the embodiments shown in the drawings. The steady acceleration detection device D in one embodiment is applied to a railcar damping device V and is used for damping a vehicle body B of the railcar. As shown in FIG. 1, the railcar damping device V includes a front actuator Af interposed between the front carriage Tf and the vehicle body B, and a rear carriage Tr and the vehicle body B. A rear actuator Ar to be mounted and a controller C that actively controls the actuators Af and Ar are provided. The steady acceleration detecting device D constitutes a part of the controller C.

  In detail, in the case of a railway vehicle, the actuators Af and Ar are connected to a pin P hanging below the vehicle body B, and are paired between the vehicle body B and the front and rear carriages Tf and Tr. It is disguised. The carriages Tf and Tr hold wheels rotatably, and a suspension spring (not shown) is interposed between the vehicle body B and the carriages Tf and Tr, and the vehicle body B is elastically supported from below. Movement of the vehicle body B in the lateral direction relative to the carriage T is allowed.

  The front and rear actuators Af and Ar basically suppress vibration in the horizontal and lateral directions with respect to the vehicle traveling direction of the vehicle body B by active control. The controller C controls the front and rear actuators Af and Ar to suppress lateral vibration of the vehicle body B.

  In this example, when the controller C performs control to suppress the vibration of the vehicle body B, the lateral acceleration αf in the horizontal direction with respect to the vehicle traveling direction of the vehicle body front portion Bf of the vehicle body B, and the vehicle body B The lateral acceleration αr in the horizontal lateral direction with respect to the vehicle traveling direction of the vehicle body rear portion Br is detected. The controller C obtains a yaw acceleration ω, which is an angular acceleration around the vehicle body center G immediately above the front and rear carts Tf, Tr, based on the lateral accelerations αf, αr, and at the horizontal lateral acceleration of the center G of the vehicle body B. A certain sway acceleration β and a steady acceleration αc are obtained. Then, the controller C obtains control forces Ff and Fr to be individually generated by the actuators Af and Ar based on the yaw acceleration ω, the sway acceleration β, and the steady acceleration αc. Further, the controller C suppresses the lateral vibration of the vehicle body B by generating thrusts with the control forces Ff and Fr for the actuators Af and Ar, respectively. The steady acceleration detector D detects the steady acceleration αc based on the yaw acceleration ω and the sway acceleration β.

  Next, a specific configuration of the actuators Af and Ar will be described. These actuators Af and Ar have the same configuration. In the figure, two actuators Af and Ar are provided for each of the carriages Tf and Tr. However, only one actuator may be provided. One controller C may be provided for each actuator Af, Ar.

  As shown in FIG. 2, the actuators Af and Ar include 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, and a cylinder A rod 4 inserted into the piston 2 and connected to the other of the piston 3, the vehicle body B and the carts Tf, Tr, a rod side chamber 5 and a piston side chamber 6 partitioned by the piston 3 in the cylinder 2 and extendable and contractible. In addition to the device Cy, the tank 7 for storing the working oil, the pump 12 that can suck up the working oil from the tank 7 and supply the working oil to the rod side chamber 5, the motor 15 that drives the pump 12, and the expansion and contraction of the cylinder device Cy. And a hydraulic circuit HC that controls thrust, and is configured as a single rod type actuator.

  In the present embodiment, 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 is provided in the middle of the first opening / closing valve 9 provided in the middle of the first passage 8 communicating the rod side chamber 5 and the piston side chamber 6 and the second passage 10 communicating the piston side chamber 6 and the tank 7. And a second on-off valve 11 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 cylinder device Cy is extended, and the second opening / closing valve 11 When the two passages 10 are in a communicating state, the first on-off valve 9 is closed and the pump 12 is driven, the cylinder device Cy contracts.

  Hereinafter, each part of the actuators Af and Ar 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.

  Note that the outer periphery of the rod guide 14 and the cylinder 2 are 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.

  Further, in the case of this cylinder device Cy, the cross-sectional area of the rod 4 is halved of the cross-sectional area of the piston 3, and the pressure receiving area on the rod side chamber 5 side of the piston 3 is half of the pressure receiving area on the piston side chamber 6 side. It is supposed to be. Therefore, if the pressure in the rod side chamber 5 is made the same during the extension operation and the contraction operation, the thrust generated in both expansion and contraction becomes equal, and the amount of hydraulic oil relative to the displacement amount of the cylinder device Cy is the same in both expansion and contraction. .

  Specifically, when the cylinder device Cy 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 actuators Af and Ar generate 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 cylinder device Cy is contracted, the rod side chamber 5 and the piston side chamber 6 are disconnected from each other, and the piston side chamber 6 is connected to the tank 7. Then, the actuators Af and Ar generate 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 actuators Af and Ar 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 thrusts of the actuators Af and Ar are controlled, the pressure in the rod side chamber 5 may be controlled for both the extension operation and the contraction operation. Further, in the actuators Af and Ar of this example, the pressure receiving area on the rod side chamber 5 side of the piston 3 is set to ½ of the pressure receiving area on the piston side chamber 6 side, so that the same thrust is generated on both expansion and contraction sides. Since the pressure in the rod side chamber 5 is the same on the expansion side and 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. 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 the expansion and contraction sides of the actuators Af and Ar is controlled by the pressure of the rod side chamber 5. The point that can be done does not change.

  Returning, the lid 13 that closes the left end of the rod 4 in FIG. 2 and the right end of the cylinder 2 is provided with a mounting portion (not shown), and the actuators Af and Ar are connected to the vehicle body B and the carriages Tf and Tr in the railway vehicle. Can be intervened between.

  The rod side chamber 5 and the piston side chamber 6 communicate with each other by 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 driven by a motor 15 that is controlled by the controller C and rotates at a predetermined rotational speed, and is a pump that discharges hydraulic oil in only 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 only discharges the hydraulic oil in one direction and does not switch the rotation direction, so there is no problem that the discharge amount changes at the time of rotation switching, and an inexpensive gear pump or the like can be used. . 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.

  Furthermore, in addition to the above-described configuration, the hydraulic circuit HC of the present example includes a discharge passage 21 that connects the rod side chamber 5 and the tank 7, and a variable relief that can change the valve opening pressure provided in the middle of the discharge passage 21. A valve 22 is provided.

  In this example, the variable relief valve 22 is a proportional electromagnetic relief valve. The variable relief valve 22 can adjust the valve opening pressure in accordance with the amount of current supplied, and when the current amount becomes maximum, the valve opening pressure is minimized, If there is no supply, the valve opening pressure is maximized.

  Thus, when the discharge passage 21 and the variable relief valve 22 are provided, the pressure in the rod side chamber 5 can be adjusted to the valve opening pressure of the variable relief valve 22 when the cylinder device Cy is expanded and contracted, and the actuators Af, Ar Can be controlled by the amount of 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 of the actuators Af and Ar are not necessary, and the motor 15 needs to be highly controlled for adjusting the discharge flow rate of the pump 12. Also disappear. Therefore, the railcar damping device V 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 actuators Af and Ar can exert 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 bogies Tf and Tr of the railway vehicle, the actuators Af and Ar are set so as not to produce the damping force in such a direction. Can function with a single-effect damper. Therefore, the actuators Af and Ar can easily function as a semi-active damper because semi-active control based on Karnop's Skyhook theory can be easily realized.

  If a proportional electromagnetic relief valve that proportionally changes the valve opening pressure with the amount of current applied to the variable relief valve 22 is used, the control of the valve opening pressure is simplified. However, any variable relief valve that can adjust the valve opening pressure is used. It is not limited to a proportional electromagnetic relief valve.

  The variable relief valve 22 has an excessive input in the expansion / contraction direction to the cylinder device Cy regardless of the open / close state of the first open / close valve 9 and the second open / close valve 11, and the pressure in the rod side chamber 5 increases the open valve pressure. When it exceeds, 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. Thus, the entire system of the actuators Af and Ar is protected. Therefore, if the discharge passage 21 and the variable relief valve 22 are provided, the system can be protected.

  Further, the hydraulic circuit HC in the actuators Af and Ar 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 flow of hydraulic oil from the tank 7 to the piston side chamber 6. A suction passage 19 that allows only air is provided. Therefore, in the actuators Af and Ar of this example, when the cylinder device Cy expands and contracts while the first on-off valve 9 and the second on-off valve 11 are closed, the 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 actuators Af and Ar in this example are 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 the two passages 10.

  In the actuators Af and Ar configured in this way, even if both the first on-off valve 9 and the second on-off valve 11 are in the shut-off position, the rod side chamber 5 and the piston in the rectifying passage 18, the suction passage 19 and the discharge passage 21. The side chamber 6 and the tank 7 are connected in a daisy chain. The rectifying passage 18, the suction passage 19, and the discharge passage 21 are set as one-way passages. Therefore, when the cylinder device Cy expands and contracts due to an external force, the hydraulic oil is always discharged from the cylinder 2 and returned to the tank 7 through the discharge passage 21, and the hydraulic oil that is insufficient in the cylinder 2 passes from the tank 7 through the suction passage 19. Supplied into the cylinder 2. Since the variable relief valve 22 acts as a resistance against the flow of the hydraulic oil to adjust the pressure in the cylinder 2 to the valve opening pressure, the actuators Af and Ar function as passive uniflow type dampers.

  Further, at the time of failure in which the actuators Af and Ar cannot be energized, each of the first on-off valve 9 and the second on-off valve 11 takes the cutoff position, and the variable relief valve 22 has a valve opening pressure. It functions as a pressure control valve fixed to the maximum. Therefore, during such a failure, the actuators Af and Ar automatically function as passive dampers.

  Subsequently, when causing the actuators Af and Ar to exert desired thrust in the extending direction, the controller C basically rotates the motor 15 to supply hydraulic oil from the pump 12 into the cylinder 2 while The on-off valve 9 is in the communication position, and the second on-off valve 11 is in the shut-off 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, and the piston 3 is pushed to the left in FIG. Demonstrate thrust. 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 amount of current applied to the variable relief valve 22. The actuators Af and Ar extend 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 direction thrust.

  On the other hand, when causing the actuators Af and Ar to exert a thrust in a desired contraction direction, the controller C rotates the motor 15 to supply hydraulic oil from the pump 12 into the rod side chamber 5, and the first on-off valve 9 is the shut-off position, and the second on-off valve 11 is 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. 2 and the actuators Af and Ar contract. Demonstrate direction thrust. As described above, when the current amount of the variable relief valve 22 is adjusted, the actuators Af and Ar cause the pressure receiving area of the piston 3 on the rod side chamber 5 side and the pressure in the rod side chamber 5 controlled by the variable relief valve 22 to be adjusted. Demonstrates the thrust in the contracted direction.

  Here, when the actuators Af and Ar do not expand / contract with an external force but expand and contract themselves, the upper limit of the pressure in the rod side chamber 5 is limited to the discharge pressure of the pump 12 driven by the motor 15. That is, when the actuators Af and Ar do not expand and contract by an external force but expand and contract by themselves, the upper limit of the pressure in the rod side chamber 5 is limited to the maximum torque that the motor 15 can output.

  In addition, the actuators Af and Ar can function not only as actuators but also as dampers 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 actuators Af and Ar 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.

  Note that the actuators Af and Ar of this example are set to a single rod type, so that it is easier to secure a stroke length than the double rod type actuator, and the total length of the actuator is shortened. Improves mounting capability.

  In addition, the hydraulic oil flow from the pump 12 in the actuators Af and Ar of this example and the flow of the hydraulic oil due to the expansion and contraction operation passes through the rod side chamber 5 and the piston side chamber 6 in order, and finally returns to the tank 7. ing. For this reason, even if gas is mixed into the rod side chamber 5 or the piston side chamber 6, the cylinder device Cy is automatically discharged to the tank 7 by the expansion / contraction operation. Therefore, when manufacturing the actuators Af and Ar, troublesome assembly in oil and assembly in a vacuum environment are not required, and advanced degassing of the hydraulic oil is not necessary, which improves productivity and manufacturing cost. Can be reduced. Further, even if gas is mixed into 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 cylinder device Cy, so that maintenance for performance recovery must be frequently performed. The maintenance labor and cost burden can be reduced.

  Subsequently, as shown in FIG. 3, the controller C calculates a front acceleration sensor 41f for detecting a lateral acceleration αf of the vehicle body front portion Bf as the vehicle body front side and a vehicle body rear side Br as the vehicle body rear side acceleration αr as the vehicle body rear side. The rear acceleration sensor 41r to be detected, the control calculation unit 44 for obtaining the control forces Ff and Fr to be output by the front and rear actuators Af and Ar, the motor 15, the first on-off valve 9, and the first on the basis of the control forces Ff and Fr The two opening-closing valve 11 and the drive part 45 which drives the variable relief valve 22 are provided.

  The front acceleration sensor 41f and the rear acceleration sensor 41r detect the lateral accelerations αf and αr as positive values when they are directed upward with respect to an axis passing through the center of the vehicle body B in FIG. On the other hand, when the direction is the downward direction in FIG. 1, it is detected as a negative value.

  Hereinafter, each part of the controller C will be described in detail. As shown in FIG. 4, the control calculation unit 44 includes a yaw suppression force calculation unit 50 that calculates a yaw suppression force fω that suppresses yaw of the vehicle body B, and a sway suppression force that calculates a sway suppression force fβ that suppresses the sway of the vehicle body B. The calculation unit 51, the centering force calculation unit 52 for obtaining the centering force fn in the direction to return the vehicle body B to the neutral position, the centering force gain change unit 53, the suppression force gain change unit 54, and the actuators Af and Ar exhibit. And a control force calculation unit 55 for obtaining power control forces Ff and Fr.

  As shown in FIG. 5, the yaw suppression force calculation unit 50 includes a yaw acceleration calculation unit 501 that obtains the yaw acceleration ω from the lateral accelerations αf and αr, a band pass filter 502 for the first straight section that filters the yaw acceleration ω, The first curve section bandpass filter 503 for filtering the yaw acceleration ω, the straight section yaw control section 504, the curved section yaw control section 505, and the straight section yaw control obtained by the straight section yaw control section 504. A gain multiplication unit 506 that multiplies the force fωs by the straight section gain Gs, a gain multiplication section 507 that multiplies the curve section yaw suppression force fωc obtained by the curve section yaw control section 505 by the curve section gain Gc, And an adding unit 508 for obtaining the yaw suppression force fω.

  The yaw acceleration calculation unit 501 divides the difference between the front lateral acceleration αf and the rear lateral acceleration αr by 2, and the yaw acceleration around the vehicle body center G immediately above each of the front cart Tf and the rear cart Tr. Find ω. The yaw acceleration calculation unit 501 obtains the yaw acceleration ω in the direction of rotating the vehicle body B in the clockwise direction around the vehicle body center G as a positive value and the yaw acceleration ω in the opposite direction as a negative value. For the convenience of obtaining the yaw acceleration ω, the installation location of the front acceleration sensor 41f may be arranged on a line along the front-rear direction or the diagonal direction including the center G of the vehicle body B and in the vicinity of the front actuator Af. Further, the installation location of the rear acceleration sensor 41r may be arranged on a line along the front-rear direction or the diagonal direction including the center G of the vehicle body B and in the vicinity of the rear actuator Ar. However, since the yaw acceleration ω can be obtained from the distance and positional relationship between the center G, the front acceleration sensor 41f and the rear acceleration sensor 41r, and the lateral accelerations αf and αr, the front acceleration sensor 41f and the rear acceleration sensor 41r are arbitrarily set. You can do it. In this case, the yaw acceleration ω is not obtained by dividing the difference between the lateral acceleration αf and the lateral acceleration αr by 2, but the difference between the lateral acceleration αf and the lateral acceleration αr, the center G of the vehicle body B, The yaw acceleration ω may be obtained from the distance and positional relationship with the acceleration sensors 41f and 41r. Specifically, when the longitudinal distance Lf between the front acceleration sensor 41f and the center G of the vehicle body B and the longitudinal distance Lr between the front acceleration sensor 41r and the center G of the vehicle body B, the yaw acceleration ω is expressed as ω = ( αf−αr) / (Lf + Lr). In this example, the yaw acceleration ω is obtained by detecting the acceleration with the front acceleration sensor 41f and the front acceleration sensor 41r, but may be detected using a yaw acceleration sensor. The yaw acceleration ω obtained by the yaw acceleration calculation unit 501 is also input to the centering force calculation unit 52.

  The band pass filter 502 for the first straight section is provided for the purpose of extracting a component of the resonance frequency band of the vehicle body B when the railway vehicle travels in the straight section at the yaw acceleration ω. Since the vehicle body B elastically supported by the carriages Tf and Tr does not normally come into contact with a stopper (not shown) that restricts the lateral movement of the vehicle body B relative to the carriages Tf and Tr to a restricted range when traveling in a straight section, The resonant frequency is between 1 Hz and 1.5 Hz. Therefore, the band pass filter 502 for the first straight section filters the yaw acceleration ω obtained by the yaw acceleration calculation unit 501 and extracts the frequency band components from 1 Hz to 1.5 Hz included in the yaw acceleration ω.

  The bandpass filter 503 for the first curve section is provided for the purpose of extracting the component of the resonance frequency band of the vehicle body B when the railway vehicle travels in the curve section at the yaw acceleration ω. The vehicle body B elastically supported by the carriages Tf and Tr is assumed to be in contact with the stopper (not shown) of the vehicle body B when traveling in a curved section, and the resonance frequency of the vehicle body B is equal to the linear section as much as the vehicle body B contacts the stopper. It is higher than when driving and is between 2 Hz and 3 Hz. Therefore, the band-pass filter 503 for the first curve section filters the yaw acceleration ω obtained by the yaw acceleration calculation unit 501 and extracts components in the frequency band from 2 Hz to 3 Hz included in the yaw acceleration ω.

  The straight section yaw control unit 504 is an H∞ controller, and the straight section yaw for suppressing the yaw of the vehicle body B from the resonance frequency band component of the yaw acceleration ω extracted by the first straight section bandpass filter 502. The suppression force fωs is calculated. The component of the resonance frequency band of the yaw acceleration ω extracted by the band pass filter 502 for the first straight section is the vibration acceleration in the resonance frequency band of the vehicle body B in the yaw direction when traveling in the straight section. Therefore, the straight section yaw suppression force fωs obtained by the straight section yaw control unit 504 is a suppression force suitable for suppressing vibration in the yaw direction of the vehicle body B during traveling in the straight section.

  The curve section yaw control unit 505 is an H∞ controller, and for the curve section that suppresses the yaw of the vehicle body B from the resonance frequency band component of the yaw acceleration ω extracted by the first curve section bandpass filter 503. The yaw suppression force fωc is calculated. The component of the resonance frequency band of the yaw acceleration ω extracted by the bandpass filter 503 for the first curve section is the vibration acceleration in the resonance frequency band of the vehicle body B in the yaw direction when traveling in the curve section. Therefore, the curve section yaw suppression force fωc obtained by the curve section yaw control unit 505 is a suppression force suitable for suppressing vibration in the yaw direction of the vehicle body B during traveling in the curve section.

  The gain multiplier 506 multiplies the linear section yaw suppression force fωs obtained by the straight section yaw control section 504 by the straight section gain Gs and outputs the result. The gain multiplication unit 507 multiplies the curve segment yaw suppression force fωc obtained by the curve segment yaw control unit 505 by the curve segment gain Gc and outputs the result.

  The straight section gain Gs and the curve section gain Gc are changed from 0 to 1 according to the determination result of the traveling section of the railway vehicle by the restraining force gain changing section 54, as shown in FIG. The straight section gain Gs takes a value of 1 when the railway vehicle is traveling in a straight section, and takes a value of 0 when the railway vehicle is traveling in a curved section. On the other hand, the curve section gain Gc takes a value of 0 when the railway vehicle is traveling in a straight section, and takes a value of 1 when the railway vehicle is traveling in a curved section. Further, the straight line gain Gs and the curved line gain Gc are changed so that the values gradually change from 0 to 1 or from 1 to 0, and the sum of both values is always 1. The total value of both is set to 1 even during the transition from 1 to 1 or from 1 to 0. The change of each of the gains Gs and Gc will be described in detail together with the explanation of the suppression force gain changing unit 54 described later.

  Then, the adding unit 508 adds the linear section yaw suppression force fωs multiplied by the straight section gain Gs and the curved section yaw suppression force fωc multiplied by the curved section gain Gc to obtain the final yaw suppression. Find the force fω. Therefore, the yaw suppression force fω is basically the straight section yaw suppression force fωs when the railway vehicle is traveling in the straight section, and the curved section yaw when the railway vehicle is traveling in the curved section. The suppression force fωc is obtained. That is, the straight section gain Gs and the curved section gain Gc select either the straight section yaw suppression force fωs suitable for the straight section or the curved section yaw suppression force fωc suitable for the curved section as the yaw suppression force fω. It is a coefficient to do. Further, when the straight section yaw suppression force fωs and the curve section yaw suppression force fωc are switched, the sum of the values of the straight section gain Gs and the curve section gain Gc is always 1, so that the yaw suppression force fω is too small. It does not become excessively large and control does not become unstable. When the yaw suppression force fω is obtained in this way, when the traveling section of the railway vehicle transitions from a straight section to a curved section, the straight section yaw suppression force fωs fades out due to changes in the gains Gs and Gc. The curve section yaw suppression force fωc fades in and switches between the two. Further, when the yaw suppression force fω is obtained in this way, when the traveling section of the railway vehicle transitions from the curve section to the straight section, the curve section yaw suppression force fωc fades out due to changes in the gains Gs and Gc. However, the straight section yaw suppression force fωs fades in and switches between the two.

  As shown in FIG. 7, the sway suppression force calculation unit 51 includes a sway acceleration calculation unit 511 that obtains the sway acceleration β from the lateral accelerations αf and αr, a band pass filter 512 for the second straight section that filters the sway acceleration β, , The second curve section bandpass filter 513 for filtering the sway acceleration β, the straight section sway control section 514, the curved section sway control section 515, and the straight section sway control section 514. A gain multiplier 516 that multiplies the suppression force fβs by the straight section gain Gs, a gain multiplier 517 that multiplies the curve section sway suppression force fβc obtained by the curve section sway control unit 515 by the curve section gain Gc, and finally And an adder 518 for obtaining a sue suppression force fβ.

  The sway acceleration calculating unit 511 obtains the sway acceleration β of the center G of the vehicle body B by dividing the sum of the lateral acceleration αf and the lateral acceleration αr by two. The sway acceleration β obtained by the sway acceleration calculation unit 511 is also input to the centering force calculation unit 52. The sway acceleration calculation unit 511 obtains a sway acceleration β in the upward direction with a positive value and an sway acceleration β in the opposite direction as a negative value with reference to an axis passing through the center of the vehicle body B in FIG.

  The band pass filter 512 for the second straight section is provided for the purpose of extracting the component of the resonance frequency band of the vehicle body B when the railway vehicle at the sway acceleration β travels in the straight section. The frequency band that the second straight section bandpass filter 512 allows to pass through is set to a frequency band from 1 Hz to 1.5 Hz, similarly to the first straight section bandpass filter 502. Therefore, the band pass filter 512 for the second straight section filters the sway acceleration β obtained by the sway acceleration calculation unit 511 and extracts the frequency band components from 1 Hz to 1.5 Hz included in the sway acceleration β.

  The bandpass filter 513 for the second curve section is provided for the purpose of extracting a component of the resonance frequency band of the vehicle body B when the railway vehicle at the sway acceleration β travels in the curve section. The frequency band that the second curve section bandpass filter 513 allows to pass through is set to a frequency band from 2 Hz to 3 Hz, similarly to the first curve section bandpass filter 503. Therefore, the bandpass filter for second curve section 513 filters the sway acceleration β obtained by the sway acceleration calculation unit 511 and extracts the frequency band components from 2 Hz to 3 Hz included in the sway acceleration β.

  The straight section sway control unit 514 is an H∞ controller, and for the straight section that suppresses the sway of the vehicle body B from the resonance frequency band component of the sway acceleration β extracted by the second straight section bandpass filter 512. The sway suppression force fβs is calculated. The component of the resonance frequency band of the sway acceleration β extracted by the band pass filter 512 for the second straight section is the vibration acceleration of the resonance frequency band in the sway direction of the vehicle body B when traveling in the straight section. Therefore, the straight section sway suppression force fβs obtained by the straight section sway control unit 514 is a suppression force suitable for suppressing vibration in the sway direction of the vehicle body B during traveling in the straight section.

  The curve section sway control unit 515 is an H∞ controller, and is used for a curve section that suppresses the sway of the vehicle body B from the resonance frequency band component of the sway acceleration β extracted by the second curve section bandpass filter 513. The sway suppression force fβc is calculated. The component of the resonance frequency band of the sway acceleration β extracted by the band-pass filter 513 for the second curve section is the vibration acceleration in the resonance frequency band of the vehicle body B in the sway direction when traveling in the curve section. Accordingly, the curve section sway suppression force fβc obtained by the curve section sway control unit 515 is a suppression force suitable for suppressing vibration in the sway direction of the vehicle body B during traveling in the curve section.

  The gain multiplication unit 516 multiplies the linear section sway suppression force fβs obtained by the linear section sway control unit 514 by the linear section gain Gs and outputs the result. The gain multiplication unit 517 multiplies the curve section sway suppression force fβc obtained by the curve section sway control unit 515 by the curve section gain Gc and outputs the result. The straight section gain Gs and the curved section gain Gc are the gains described above, and are gains whose values change from 0 to 1 as described above.

  Then, the adding unit 518 adds the straight section sway suppression force fβs multiplied by the straight section gain Gs and the curved section sway suppression force fβc multiplied by the curved section gain Gc to obtain a final sway suppression. Determine the force fβ. Therefore, the sway suppression force fβ basically becomes the sway suppression force fβs for the straight section when the railway vehicle is traveling in the straight section, and the sway suppression force fβs when the railway vehicle is traveling in the curved section. The suppression force fβc is obtained. That is, in the sway suppression force calculation unit 51, the straight section gain Gs and the curve section gain Gc are either the straight section sway suppression force fβs suitable for the straight section or the curved section sway suppression force fβc suitable for the curved section. Is a coefficient for selecting as the sway suppression force fβ. Further, when the straight section sway suppression force fβs and the curve section sway suppression force fβc are switched, the sum of the values of the straight section gain Gs and the curve section gain Gc is always 1, so that the sway suppression force fβ is too small. It does not become excessively large and control does not become unstable. When the sway restraining force fβ is obtained in this way, when the traveling section of the railway vehicle transitions from the straight section to the curved section, the sway restraining force fβs for the straight section fades out due to changes in the gains Gs and Gc. The sway restraining force fβc for the curve section fades in and switches between the two. In addition, when the sway suppression force fβ is obtained in this way, when the traveling section of the railway vehicle transitions from the curve section to the straight section, the curve section sway suppression force fβc fades out due to changes in the gains Gs and Gc. However, the sway restraining force fβs for the straight section fades in, and both are switched.

  As shown in FIG. 8, the centering force calculator 52 filters the yaw acceleration low-pass filter 521 that filters the yaw acceleration ω output from the yaw acceleration calculator 501 and the sway acceleration β output from the sway acceleration calculator 511. A low-pass filter for sway acceleration 522, a high-pass filter for yaw acceleration that filters the yaw acceleration ω filtered by the low-pass filter for yaw acceleration 521, and a high-pass for sway acceleration that filters the sway acceleration β filtered by the low-pass filter for sway acceleration 522 A filter 524, a steady acceleration calculation unit 525 for obtaining the steady acceleration αc by subtracting the filtered yaw acceleration ω ′ from the filtered post-filtered sway acceleration β ′, and a centering for obtaining the centering force fn from the obtained steady acceleration αc. Force calculation unit 526 and centering force fn And a gain multiplying unit 527 for multiplying the centering force gain Gn.

  The yaw acceleration low-pass filter 521 extracts the low frequency component of the yaw acceleration ω by filtering the yaw acceleration ω obtained by the yaw acceleration calculation unit 501. Specifically, the cutoff frequency of the yaw acceleration low-pass filter 521 is set to about 0.3 Hz, and a low frequency component of 0.3 Hz or less included in the yaw acceleration ω can be extracted. The filtered yaw acceleration ω 'filtered by the yaw acceleration low-pass filter 521 is filtered by the yaw acceleration high-pass filter 523, and the drift components of the acceleration sensors 41f and 41r are removed from the filtered yaw acceleration ω'. The cut-off frequency of the yaw acceleration high-pass filter 523 is set to a value sufficiently lower than the cut-off frequency of the yaw acceleration low-pass filter 521, and is set to a frequency that does not affect the obtaining of the steady acceleration αc. The

  The low-pass filter for sway acceleration 522 filters the sway acceleration β obtained by the sway acceleration calculation unit 511 and extracts a low-frequency component included in the sway acceleration β. Specifically, the cutoff frequency of the low-pass filter for sway acceleration 522 is set to about 0.3 Hz, and a component of 0.3 Hz or less included in the sway acceleration β can be extracted. The filtered sway acceleration β 'filtered by the sway acceleration low-pass filter 522 is filtered by the sway acceleration high-pass filter 524, and the drift components of the acceleration sensors 41f and 41r are removed from the filtered sway acceleration β'. The cutoff frequency of the high-pass filter for sway acceleration 524 is set to a value sufficiently lower than the cutoff frequency of the low-pass filter for sway acceleration 522, and is set to a frequency that does not affect the steady acceleration αc. The

  The steady acceleration calculation unit 525 obtains a steady acceleration αc. Basically, the low-pass filter 522 for sway acceleration is used to extract the steady acceleration component included in the sway acceleration β. However, when the sway acceleration β is filtered by the low-pass filter, the phase of the sway acceleration β ′ after filtering becomes the sway acceleration β. Will be delayed with respect to the phase. Therefore, since the post-filtered sway acceleration β ′ subjected to the low-pass filter processing is delayed in phase from the actual steady-state acceleration αc, if the post-filtered sway acceleration β ′ subjected to the low-pass filter processing is simply set as the steady acceleration, the steady acceleration can be accurately obtained. I can't get it.

  When paying attention to the yaw acceleration ω, there are the following properties. When a railway vehicle enters a curved section (including a relaxation curved section installed before the curved section; the same applies hereinafter), the front carriage Tf moves relative to the vehicle body B toward the center of curvature. Therefore, in a scene where the railway vehicle enters the curved section from the straight section, the suspension spring is bent by the movement of the carriage of the front carriage Tf, and a yaw acceleration ω is generated that pushes the front side of the vehicle body B toward the center of curvature. When the railway vehicle enters the straight section from the curved section, the front carriage Tf with respect to the vehicle body B moves toward the center of curvature. Therefore, in a scene where the railway vehicle enters from the curved section to the straight section, the suspension spring is bent by the movement of the carriage of the front carriage Tf, and the yaw acceleration ω that pushes the front side of the vehicle body B toward the center of curvature is generated. On the other hand, the post-filtering sway acceleration β ′ filtered by the low-pass filter for sway acceleration 522 is an acceleration equivalent to a steady acceleration acting on the vehicle body B when the railway vehicle travels in a curved section, and the vehicle body B is on the side of the center of curvature. However, the phase is delayed with respect to the actual steady acceleration αc as described above. Since the carriage Tf on the front side of the vehicle body B first moves along the track at the entrance of the curved section, a yaw acceleration ω opposite to the sway acceleration β is generated at a timing earlier than the sway acceleration β. Since the carriage Tf on the front side of the vehicle body B first moves along the track at the exit of the curved section, the yaw acceleration ω in the same direction as the sway acceleration β is generated at a timing earlier than the value of the sway acceleration β decreases.

  From the above, if the steady acceleration αc is obtained by subtracting the filtered yaw acceleration ω ′ from the filtered sway acceleration β filtered by the low-pass filter for sway acceleration 522, there is no phase lag or a phase lag with respect to the actual steady acceleration. A small steady acceleration αc is obtained. Since the yaw acceleration ω is filtered by the yaw acceleration low-pass filter 521, the obtained steady acceleration αc does not have a waveform whose phase advances with respect to the actual steady acceleration. Therefore, the steady acceleration calculation unit 525 obtains the steady acceleration αc by subtracting the filtered yaw acceleration ω from the filtered sway acceleration β. Thus, if steady acceleration (alpha) c is calculated | required, the steady acceleration at the time of a rail vehicle drive | working a curve area can be calculated | required accurately. From the above, the steady acceleration detection device D is, in this example, the yaw acceleration detection unit Y including the front acceleration sensor 41f, the rear acceleration sensor 41r, and the yaw acceleration calculation unit 501, the front acceleration sensor 41f, and the rear acceleration sensor. 41r and a sway acceleration calculating unit 510, a sway acceleration detecting unit S, a yaw acceleration low-pass filter 521 for filtering the yaw acceleration ω, and a sway acceleration low-pass filter 522 for filtering the sway acceleration β output by the sway acceleration calculating unit 511 The stationary acceleration calculating unit 525 calculates the steady acceleration αc by subtracting the filtered yaw acceleration ω ′ from the filtered sway acceleration β ′. If there is no problem of drift of the acceleration sensors 41f and 41r, the high-pass filter for yaw acceleration 523 and the high-pass filter 524 for sway acceleration may be omitted. In addition, the yaw acceleration detection unit that detects the yaw acceleration ω may be a yaw acceleration sensor that detects the yaw acceleration ω of the vehicle body B. The yaw acceleration detector Y and the sway acceleration detector S may be incorporated in an external device other than the railcar vibration damping device V.

  The centering force calculation unit 526 obtains the centering force fn from the steady acceleration αc. Here, the maximum value of the steady acceleration αc allowed when the railway vehicle travels in the curved section is αcmax, and the maximum value of the force that the actuators Af and Ar can output when the motor 15 drives the pump 12 with the rated torque. Is ftmax. Then, the centering force calculation unit 526 calculates the centering force fn from the steady acceleration αc by calculating the following equation: fn = αc × ftmax / αcmax. If the steady acceleration αc exceeds αcmax, the value of the steady acceleration αc is limited to αcmax. Therefore, the upper limit of the centering force fn is set to the maximum value of the force that the actuators Af and Ar can exert when the motor 15 drives the pump 12 with the rated torque. The maximum value αcmax of the steady acceleration αc is a predetermined value.

  The gain multiplier 527 multiplies the centering force fn by the centering force gain Gn and outputs the result. The centering force gain Gn is changed from 0 to 1 by the centering force gain changing unit 53.

  The centering force gain changing unit 53 changes the centering force gain Gn based on the steady acceleration αc detected by the steady acceleration detecting device D. Specifically, the centering force gain changing unit 53 performs absolute value processing on the steady acceleration αc output from the steady acceleration calculating unit 525, and compares the absolute value of the steady acceleration αc with the centering threshold value α1. Then, as a result of the comparison, the centering force gain changing unit 53 sets the centering force gain Gn to 1 and the absolute value of the steady acceleration αc is less than the centering threshold α1 when the absolute value of the steady acceleration αc is greater than or equal to the centering threshold α1. And the centering force gain Gn is set to zero.

  The centering force gain changing unit 53 gradually increases the centering force gain Gn from 0 to 1 over time when the absolute value of the steady acceleration αc increases across the value of the centering threshold α1. change. Further, the centering force gain changing unit 53 gradually decreases the centering force gain Gn from 1 to 0 with the passage of time when the absolute value of the steady acceleration αc decreases across the value of the centering threshold α1. change. As described above, since the centering force gain changing unit 53 changes the value of the centering force gain Gn as described above, the centering force fn is faded into the final control forces Ff and Fr when the centering force fn is required. Can be. Further, when the centering force fn is unnecessary, the centering force fn can be faded out from the final control forces Ff and Fr.

  The restraining force gain changing unit 54 changes the values of the straight section gain Gs and the curve section gain Gc described above based on the steady acceleration αc detected by the steady acceleration detector D. Specifically, the suppression force gain changing unit 54 performs absolute value processing on the steady acceleration αc output from the steady acceleration calculating unit 525, and compares the absolute value of the steady acceleration αc with the curve determination threshold value α2. When the absolute value of the steady acceleration αc is equal to or greater than the curve determination threshold value α2 as a result of the comparison, the suppression force gain changing unit 54 determines that the traveling section of the railway vehicle is a curved section. On the contrary, when the absolute value of the steady acceleration αc is less than the curve determination threshold value α2, the restraining force gain changing unit 54 determines that the traveling section of the railway vehicle is a straight section. As described above, the suppression force gain changing unit 54 also functions as a curve section determination unit that determines whether or not the section in which the railway vehicle is traveling is a curved section.

  As shown in FIG. 6, when the suppression force gain changing unit 54 determines that the railway vehicle is traveling in the straight section, the value of the straight section gain Gs is set to 1, and the railway vehicle is traveling in the curved section. Is determined, the value of the straight section gain Gs is set to zero. Further, as shown in FIG. 6, when the suppression force gain changing unit 54 determines that the railway vehicle is traveling in a straight section, the value of the curve section gain Gc is set to 0, and the railway vehicle is traveling in the curved section. If it is determined that the curve section gain Gc is 1, the curve section gain Gc is set to 1. The value of the straight section gain Gs is gradually changed from 1 to 0 with the passage of time when the traveling section of the railway vehicle is switched from the straight section to the curved section. Although not shown, the value of the straight section gain Gs is gradually increased from 0 to 1 when the traveling section of the railway vehicle is switched from the curved section to the straight section. The value of the curve section gain Gc is gradually increased from 0 to 1 with the passage of time when the traveling section of the railway vehicle is switched from the straight section to the curved section. Although not shown, the value of the curve section gain Gc is gradually decreased from 1 to 0 as time passes, when the traveling section of the railway vehicle is switched from the curved section to the straight section. The curve determination threshold value α2 is set to a value larger than the centering threshold value α1. The sum of the values of the straight section gain Gs and the curved section gain Gc is always 1, and the total value of both is set to 1 even when changing from 0 to 1 or from 1 to 0. Is done. The time required for the change in the values of both gains Gs and Gc can be arbitrarily set.

  As described above, the adding unit 508 in the yaw suppression force calculation unit 50 multiplies the curve segment yaw suppression force fωc by the curve segment yaw suppression force fωs to the value obtained by multiplying the linear segment yaw suppression force fωs by the curve segment gain Gc. The final yaw suppression force fω is obtained by adding the multiplied values. Therefore, when the traveling section of the railway vehicle transitions from the straight section to the curved section, the straight section yaw suppression force fωs fades out and the curved section yaw suppression force fωc fades in due to changes in the gains Gs and Gc. Switch between the two. Further, when the traveling section of the railway vehicle transitions from the curved section to the straight section, the curved section yaw suppression force fωc fades out and the straight section yaw suppression force fωs fades in due to changes in the gains Gs and Gc. Switch between the two.

  As described above, the adding unit 518 in the sway suppression force calculation unit 51 multiplies the curve segment sway suppression force fβc by the curve segment sway suppression force fβs to the curve segment sway suppression force fβs. The final value of the sway suppression force fβ is obtained by adding the multiplied values. Therefore, when the traveling section of the railway vehicle transitions from the straight section to the curved section, the straight section sway suppression force fβs fades out while the straight section sway suppression force fβs fades out due to changes in the gains Gs and Gc. Switch between the two. Further, when the traveling section of the railway vehicle transitions from the curved section to the straight section, the straight section sway suppressing force fβs fades in while the curved section sway suppressing force fβc fades out due to a change in the gains Gs and Gc. Switch between the two.

  As shown in FIG. 9, the control force calculating unit 55 determines the front actuator Af and the rear actuator from the yaw suppression force fω, the sway suppression force fβ, and the value fn · Gn obtained by multiplying the centering force fn by the centering force gain Gn. A control force calculation unit 551 for obtaining the control forces Ff and Fr of the actuator Ar and a limiter 552 are provided.

  The control force calculation unit 551 calculates a suppression force ff of the actuator Af on the front side by dividing the value obtained by adding the yaw suppression force fω and the sway suppression force fβ by 2, and adds the centering force gain Gn to the suppression force ff and the centering force fn. The control force Ff of the front actuator Af is obtained by adding the value fn · Gn multiplied by. Further, the control force calculation unit 551 obtains the suppression force fr of the rear actuator Ar by dividing the value obtained by subtracting the yaw suppression force fω from the sway suppression force fβ by 2, and the centering force fn is added to the centering force fn. The value fn · Gn multiplied by the gain Gn is added to determine the control force Fr of the rear actuator Ar. Furthermore, when the limiter 552 exceeds the upper limit by the limiter 552, the control forces Ff and Fr are limited to the upper limit value and input to the drive unit 45.

  When the control forces Ff and Fr are obtained in this way, the centering force fn fades into the control forces Ff and Fr when the traveling section of the railway vehicle transitions from the straight section to the curved section. Further, regarding the suppression forces ff and fr of the breakdown of the control forces Ff and Fr, the yaw suppression force fωc for the curve section suitable for the curve section from the yaw suppression force fωs for the straight section suitable for the straight section and the sway suppression force fβs. And the sway suppression force fβc. Since the curve determination threshold value α2 is set to a value larger than the centering threshold value α1, when the traveling section of the railway vehicle transitions from a straight section to a curved section, the restraining forces ff and fr are restraining forces suitable for the straight section. The centering force fn fades in to the control forces Ff and Fr before switching to the suppression force suitable for the curve section. Therefore, the centering force Fn is immediately exerted when the railway vehicle reaches the curved section, so that the sway of the vehicle body B can be suppressed, and the situation in which the vehicle body B compresses the stopper (not shown) most effectively can be prevented. Further, it has been found that, at the entrance of the curved section, the ride comfort is better when the yaw suppression force fωs and the sway suppression force fβs for the linear section are exerted on the actuators Af and Ar. Since the curve determination threshold value α2 is set larger than the centering threshold value α1 and is set to a value at which it is possible to determine that the travel section of the railway vehicle is completely a curve section, the yaw suppression force fωs and the sway suppression force fβs for the straight section are curved. It can be demonstrated at the entrance of the section, improving riding comfort.

  Further, when the control forces Ff and Fr are obtained as described above, the centering force fn fades out from the control forces Ff and Fr when the traveling section of the railway vehicle transitions from the curved section to the straight section. In addition, regarding the suppression forces ff and fr of the breakdown of the control forces Ff and Fr, the yaw suppression force fωs for the straight section suitable for the straight section from the yaw suppression force fωc and the sway suppression force fβc for the curve section suitable for the curved section. And the sway suppression force fβs. As described above, since the curve determination threshold value α2 is set to a value larger than the centering threshold value α1, when the traveling section of the railway vehicle transitions from the curved section to the straight section, the restraining forces ff and fr are suitable for the curved section. The centering force fn fades out from the control forces Ff and Fr after switching from the suppression force to the suppression force suitable for the straight section. Therefore, the centering force fn continues to be exerted until the railway vehicle completely enters the straight section, so that the sway of the vehicle body B can be suppressed, and the situation in which the vehicle body B compresses the stopper (not shown) most effectively can be prevented. . Further, it has been found that, at the exit of the curved section, the ride comfort is better when the yaw suppression force fωs and the sway suppression force fβs for the linear section are exerted on the actuators Af and Ar. Since the curve determination threshold value α2 is larger than the centering threshold value α1, it is easy to determine that the travel section of the railway vehicle has left the curve section, and the yaw suppression force fωs and the sway suppression force fβs for the straight section are output at the exit of the curve section. Can demonstrate. Therefore, riding comfort can be improved regardless of the travel section.

  The drive unit 45 includes a driver circuit that drives the motor 15, the first on-off valve 9, the second on-off valve 11, and the variable relief valve 22. The drive unit 45 controls the amount of current supplied to the motor 15, the first on-off valve 9, the second on-off valve 11, and the variable relief valve 22 in each actuator Af, Ar according to the control forces Ff, Fr. The actuators Af and Ar are caused to exert thrust according to the forces Ff and Fr.

  In controlling the motor 15, the drive unit 45 controls the motor 15 to rotate the motor 15 at a constant rotation speed at a predetermined rotation speed. The motor 15 can output a torque that exceeds the rated torque within a range that does not burn out. Therefore, even if the control forces Ff and Fr are values that cause the motor 15 to output a torque exceeding the rated torque, the motor 15 can output a torque exceeding the rated torque within a range that does not burn.

  Although not shown, the controller C specifically includes, for example, an A / D converter for capturing signals output from the front acceleration sensor 41f and the rear acceleration sensor 41r, and a lateral acceleration. A storage device such as a ROM (Read Only Memory) that stores programs used for processing necessary to control the actuators Af and Ar by taking in αf and lateral acceleration αr, and executes processing based on the programs And a storage device such as a RAM (Random Access Memory) that provides a storage area for the CPU. And the structure of each part of the controller C is realizable by execution of the program for performing the said process of CPU.

  The process of the controller C will be described using the flowchart shown in FIG. First, the controller C takes in the lateral acceleration αf and the lateral acceleration αr (step F1). Subsequently, the controller C obtains the yaw acceleration ω and the sway acceleration β (step F2). Further, the controller C obtains a straight section yaw suppression force fωs, a curved section yaw suppression force fωc, a straight section swage suppression force fβs, and a curved section sway suppression force fβc from the yaw acceleration ω and the sway acceleration β (step) F3). Next, the controller C obtains a steady acceleration αc from the filtered yaw acceleration ω ′ and the filtered sway acceleration β ′ (step F4). Then, the controller C obtains the centering force fn from the steady acceleration αc (step F5). Further, the controller C determines the value of the centering force gain Gn by comparing the absolute value of the steady acceleration αc and the centering threshold value α1, and further compares the absolute value of the steady acceleration αc and the curve determination threshold value α2 by the railway vehicle. It is determined whether the vehicle is traveling in a straight section but in a curved section, and the values of the straight section gain Gs and the curved section gain Gc are determined (step F6). The controller C then calculates the linear section gain Gs and the curved section gain Gc, the straight section yaw suppression force fωs, the curved section yaw suppression force fωc, the straight section sway suppression force fβs, and the curved section sway suppression force fβc. The yaw suppression force fω and the sway suppression force fβ are obtained (step F7). Subsequently, the controller C obtains a value fn · Gn obtained by multiplying the centering force fn by the centering force gain Gn (step F8). Further, the controller C obtains the control forces Ff and Fr of the front and rear actuators Af and Ar from the value fn · Gn obtained by multiplying the yaw suppression force fω, the sway suppression force fβ and the centering force fn by the centering force gain Gn (step F9). ). Finally, the controller C drives the motors 15 of the actuators Af and Ar, the first on-off valve 9, the second on-off valve 11 and the variable relief valve 22 based on the control forces Ff and Fr, and applies the actuators Af and Ar to the actuators Af and Ar. The thrust is exhibited (step F10).

  As described above, the railcar vibration damping device V includes the actuators Af and Ar that are interposed between the vehicle body B of the railway vehicle and the carriages Tf and Tr and can exert control force, and the yaw acceleration ω of the vehicle body B. A controller C that obtains control forces Ff and Fr that suppress vibrations of the vehicle body B based on the sway acceleration β. When the absolute value of the steady acceleration αc is equal to or greater than the centering threshold α1, the controller C calculates the yaw acceleration ω and the sway acceleration β. The control force Ff, based on the obtained suppression force ff, fr for suppressing the vibration of the vehicle body B in the yaw direction and the sway direction, and the centering force fn in the direction for returning the vehicle body B to the neutral position obtained based on the steady acceleration αc. Fr is obtained.

  In the railway vehicle damping device V configured as described above, when the absolute value of the steady acceleration αc detected by the steady acceleration detector D is equal to or greater than the centering threshold α1, the centering force fn obtained based on the steady acceleration αc and the suppression are suppressed. Control forces Ff and Fr are obtained based on the forces ff and fr. Therefore, the railcar damping device V determines whether the centering force fn needs to be exhibited based on the value of the steady acceleration αc and does not require a displacement sensor. According to the railcar vibration damping device V of the present invention, the restraining forces ff and fr and the centering force fn for suppressing vibration during traveling in a curved section can be exerted, and the vehicle body B is brought into contact with the stopper for maximum compression. Since it can suppress, it can suppress that the vibration from the trolley | bogie Tf and Tr side transmits to the vehicle body B at the time of curve area travel. As described above, according to the railcar damping device V of the present invention, it is possible to improve the riding comfort during traveling in a curved section.

  Here, the steady acceleration αc acting on the vehicle body B traveling in the curved section is a component included in the sway acceleration β, but when trying to obtain the steady acceleration by the low-pass filter process for extracting the low frequency component of the sway acceleration β. Actually, the phase is delayed from the steady acceleration. On the other hand, in the steady acceleration detecting device D of the present invention, the filtered sway acceleration β ′ obtained by processing the sway acceleration β by the low-pass filter 522 for sway acceleration and the yaw acceleration ω processed by the low-pass filter 521 for yaw acceleration. The steady acceleration αc is obtained based on the rear yaw acceleration ω ′. At the entrance and exit of the curved section, the yaw acceleration ω acts on the vehicle body B at a timing earlier than the low-frequency component of the sway acceleration β obtained by filtering with the sway acceleration low-pass filter 522. Therefore, if not only the sway acceleration β obtained by filtering with the sway acceleration low-pass filter 522 but also the yaw acceleration ω is added, the phase lag is compensated and there is no or very little phase lag with respect to the actual steady acceleration. There is no or very little steady acceleration αc. Therefore, according to the steady acceleration detector D of the present invention, the steady acceleration αc can be detected with high accuracy.

  In the steady acceleration detecting device D, specifically, the steady acceleration αc is obtained by subtracting the filtered yaw acceleration ω ′ from the filtered sway acceleration β ′. The yaw acceleration ω acts on the front side of the vehicle body B in the opposite direction to the sway acceleration β at the entrance of the curve section, and acts on the front side of the vehicle body B in the same direction as the sway acceleration β at the exit of the curve section. Therefore, when the yaw acceleration ω is subtracted from the sway acceleration β, the phase delay of the obtained steady acceleration αc with respect to the actual steady acceleration can be eliminated using a simple calculation.

  A value obtained by multiplying the yaw acceleration ω by a coefficient may be subtracted from the sway acceleration β. The coefficient may be set so as to be optimal according to the standard and behavior of the railway vehicle.

  Instead of subtracting the filtered yaw acceleration ω ′ from the filtered sway acceleration β ′, the sway acceleration β is multiplied by a gain that changes according to the value of the yaw acceleration ω, and the phase is delayed from the actual steady acceleration. The steady acceleration αc that does not become may be obtained. Since the sway acceleration β is smaller than the actual steady acceleration at the entrance of the curve section and the sway acceleration β is larger than the actual steady acceleration at the exit of the curve section, the gain varies depending on the value of the filtered yaw acceleration ω ′. Can be obtained by multiplying the sway acceleration β ′ after filtering by the steady acceleration αc in which the phase delay is eliminated. When changing the gain, the value of the yaw acceleration ω and the gain value may be mapped, and the gain value may be obtained by map calculation.

  In the railcar vibration damping device V, the steady acceleration αc detected by the steady acceleration detecting device D is used to determine whether or not the section in which the railway vehicle is traveling is a curved section. Curve section determination can be performed accurately without using travel point information from a monitor or GPS, or without using a special sensor for detecting the travel position.

  Further, in the railcar vibration damping device V, the centering force fn for returning the vehicle body B to the neutral position is obtained based on the steady acceleration αc with little or no phase delay. A centering force fn that suppresses only eccentricity from the position is obtained. Therefore, the centering force fn without excess or deficiency can be exhibited, and the eccentricity of the vehicle body B can be effectively suppressed.

  Further, in the railcar damping device V, displacement feedback control is not performed, and it is possible to suppress transmission of vibrations from the carts Tf and Tr to the vehicle body B without interfering with control for suppressing vibration of the vehicle body B. . As described above, in the railcar vibration damping device V of the present invention, a displacement sensor is not necessary for determination of travel in a curved section, and displacement feedback control that impedes riding comfort is not performed, and centering is performed based on steady acceleration αc. Since the force fn is obtained, it is possible to improve riding comfort when traveling in a curved section. Therefore, according to the railcar vibration damping device V of the present invention, the displacement sensor is not required, the cost can be reduced, and the riding comfort during traveling in the curved section can be improved.

  Although it is possible to determine whether the railway vehicle is traveling in a curved section from the point information available from the vehicle monitor mounted on the railway vehicle, there is an error in the point information and the centering is not performed. The force fn may be exerted. On the other hand, in the railcar vibration damping device V of the present invention, the steady acceleration αc with no or very little phase delay is compared with the centering threshold α1, and when the steady acceleration αc is equal to or greater than the centering threshold α1, the centering force fn is obtained. In order to output, since the centering force fn is immediately exerted when the railway vehicle reaches the curved section, the most contraction of the stopper can be effectively suppressed and the riding comfort can be further improved. The centering threshold value α1 may be a value of the steady acceleration αc detected when the railway vehicle approaches the curved section.

  In the railcar damping device V of this example, the centering force fn is set to the upper limit of the centering force fn as the maximum value of the force that the actuators Af and Ar can exert when the motor 15 drives the pump 12 with the rated torque. It comes to ask for. In the railway vehicle vibration damping device V configured in this way, even if the actuators Af and Ar output only the centering force fn, the remaining force remains until the maximum torque that the motor 15 can output. The suppression forces ff and fr for suppressing the vibration of the vehicle body B can be output while exerting the force fn. Therefore, according to the railcar vibration damping device V2 of the present example, it is possible to exhibit the restraining forces ff and fr that suppress the vibration of the vehicle body B while exhibiting the centering force fn that returns the vehicle body B to the neutral position when traveling in a curved section. The riding comfort during traveling in a curved section can be further improved.

  Further, in the railway vehicle vibration damping device V of this example, in order to obtain the suppression forces ff and fr, the straight section yaw control section 504 and the straight section sway control section 514 as the straight section control section, and the curved section section A curve section yaw control section 505 and a curve section sway control section 515 are provided as control sections. When the absolute values of the steady-state acceleration αc are greater than the centering threshold value α1 and the curve determination threshold value α2 is greater than the curve determination threshold value α2 and is greater than or equal to the curve determination threshold value α2, the restraining forces ff and fr are controlled for the straight section. When the absolute value of the steady acceleration αc is greater than or equal to the curve determination threshold value α2 and less than the curve determination threshold value α2 is switched from the suppression force ff and fr required by the control unit to the suppression force ff and fr required by the curve section control unit. The switching force ff, fr required by the unit is switched to the suppression force ff, fr required by the linear section control unit. The straight section control unit obtains suppression forces ff and fr suitable for suppressing the lateral vibration of the vehicle body B when traveling in the straight section, and the curved section control unit calculates the lateral vibration of the vehicle body B when traveling in the curved section. Suppression forces ff and fr suitable for suppression of the above are obtained. Therefore, according to the railcar damping device V of the present example, the optimum control forces Ff and Fr can be exhibited according to the traveling section of the railway vehicle, so that a high vibration suppressing effect can be obtained regardless of the traveling section. Further, whether or not the centering force fn can be output is based on the centering threshold α1, and the switching between the control for the straight section and the control for the curved section is based on the curve determination threshold α2 that is larger than the centering threshold α1. Therefore, the curve determination threshold value α2 can be set larger than the centering threshold value α1, and can be set to a value at which it is possible to determine that the traveling section of the railway vehicle is completely a curved section. The yaw suppression force fωs and the sway suppression force fβs for the straight section It can be demonstrated at the entrance and exit of the car, improving ride comfort.

  It is preferable to set the curve determination threshold α2 to a value larger than the centering threshold α1, but it is also possible to set both to the same value. In this case, the gain multiplier 527 uses the curve section instead of the centering force gain Gn. The centering gain force changing unit 53 may be omitted by multiplying the gain Gc for use by the centering force fn.

  Further, in the railcar vibration damping device V of this example, the switching force ff, fr obtained by the straight section control unit and the restraining force ff, fr obtained by the curved section control unit are selected before switching. The restraining forces ff and fr that have been applied are faded out, and the restraining forces ff and fr that should be selected after switching are faded in. According to the railcar damping device V configured as described above, the values of the restraining forces ff and fr change suddenly when the restraining forces ff and fr for the straight section and the restraining forces ff and fr for the curved section are switched. Therefore, stability in control is improved. Further, when the linear section gain and curve section fading-in and fade-out of the restraining forces ff and fr for the straight section and the straight section gain Gs and the curved section gain Gc are used, and the sum of the two is always 1, the final restraining forces ff and fr Does not become too small or too large, and control does not become unstable.

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

12 ... Pump, 15 ... Motor, 504 ... Straight section yaw control section (straight section control section), 505 ... Curve section yaw control section (curve section control section), 514 ..Sway control section for straight section (control section for straight section) 515... Sway control section for curved section (control section for curved section) 521... Low pass filter for yaw acceleration 522. Low-pass filter, 525 ... Steady acceleration calculation unit, Af, Ar ... Actuator, B ... Car body, C ... Controller, Cy ... Cylinder device, D ... Steady acceleration detection device, S ... Sway acceleration detection unit, Tf, Tr ... cart, V ... Railway vehicle vibration damping device, Y ... Yaw acceleration detection unit

Claims (6)

A sway acceleration detector that detects the sway acceleration of the body of the railway vehicle;
A yaw acceleration detector for detecting the yaw acceleration of the vehicle body;
A low-pass filter for sway acceleration that filters the sway acceleration detected by the sway acceleration detector;
A low-pass filter for yaw acceleration that filters the yaw acceleration detected by the yaw acceleration detector,
Steady state for obtaining a steady acceleration when the railway vehicle travels a curved section based on the filtered sway acceleration processed by the low-pass filter for sway acceleration and the post-filtered yaw acceleration processed by the low-pass filter for yaw acceleration A steady-state acceleration detection device comprising an acceleration calculation unit. The steady acceleration detection device according to claim 1, wherein the steady acceleration calculation unit obtains the steady acceleration by subtracting the filtered yaw acceleration from the filtered sway acceleration. An actuator that is interposed between the body of the railway vehicle and the bogie and can exert control power;
A controller for controlling the actuator by obtaining the control force for suppressing vibration of the vehicle body based on the yaw acceleration and the sway acceleration of the vehicle body,
The steady acceleration detection device according to claim 1 or 2,
When the absolute value of the steady acceleration is equal to or greater than a centering threshold, the controller includes a suppression force that suppresses vibrations in the yaw direction and the sway direction of the vehicle body that are obtained based on the yaw acceleration and the sway acceleration, and the steady acceleration. The vibration control device for a railway vehicle, characterized in that the control force is obtained based on a centering force in a direction to return the vehicle body to a neutral position. The actuator includes a cylinder device that can be expanded and contracted, and a pump that is driven by a motor and can supply the working liquid to the cylinder device, and exhibits the control force by supplying the working liquid to the cylinder device. ,
The said controller calculates | requires the said centering force by making the upper limit of the said centering force into the maximum value of the force which the said actuator can exhibit when the said motor drives the said pump with a rated torque. Railway vehicle vibration control device. The controller is
It has a control section for a straight section and a control section for a curve section for obtaining the suppression force,
The suppression force is based on a curve determination threshold value that is larger than the centering threshold value, and when the absolute value of the steady acceleration is less than the curve determination threshold value or more than the curve determination threshold value, the curve section is calculated based on the suppression force obtained by the linear section control unit. It is switched to the suppression force required by the section control unit,
The suppression force is based on a curve determination threshold value that is larger than the centering threshold value, and when the steady acceleration is greater than or equal to the curve determination threshold value and less than the curve determination threshold value, the suppression force obtained by the curve interval control unit is used for the straight line interval. The vibration control device for a railway vehicle according to claim 3 or 4, wherein the control unit is switched to a suppression force required by the controller. The controller fades out the suppression force selected before switching when switching between the suppression force calculated by the linear section control unit and the suppression force calculated by the curve section control unit, and should be selected after switching. The railway vehicle vibration damping device according to claim 5, wherein the force is faded in.

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