JP2024108368A - Vibration control devices for railway vehicles - Google Patents

Abstract

A vibration damping device for railway vehicles is provided that is capable of generating the necessary damping force required when an earthquake occurs, while at the same time reducing production costs.
[Solution] The vibration control device S for railway vehicles of the present invention has a damper D and either a semi-active damper SD or an actuator A interposed in parallel between the carbody B and bogie T of the railway vehicle, and the required damping force is the sum of the damping forces generated by the damper D and either the semi-active damper SD or the actuator A required when expanding and contracting at a predetermined stroke speed expected in the event of an earthquake, and the damper D generates a damping force greater than the required damping force when it expands and contracts at a stroke speed faster than the predetermined stroke speed.
[Selected figure] Figure 2

Description

The present invention relates to a vibration damping device for railway vehicles.

Conventionally, this type of damper is used as a railway vehicle vibration control device that is interposed between the car body and bogie of a railway vehicle and suppresses vibrations in the left-right direction relative to the direction of travel of the car body. In addition to the damper, the railway vehicle vibration control device also includes an actuator, and the force exerted by the damper and actuator suppresses vibrations of the car body.

However, if a strong earthquake occurs while a railway vehicle is running, the car body may shake violently and may derail. In such a case, it is desirable for the railway vehicle vibration control device to provide a high damping force using dampers and actuators to prevent derailment.

In conventional railway vehicle vibration control devices, the required damping force is the sum of the damping forces required to be generated by the damper and actuator when an earthquake occurs, and the required damping force is shared equally between the damper and actuator when an earthquake occurs.

JP 2019-043297 A

As described above, in conventional railway vehicle vibration control devices, the damper and actuator equally shared the required damping force during an earthquake, so the damper and actuator had to be designed to withstand high pressure. In addition, valves to be used during an earthquake had to be provided in both the damper and the actuator, complicating the structure and increasing manufacturing costs.

Therefore, the present invention aims to provide a vibration damping device for railway vehicles that can reduce manufacturing costs while still generating the necessary damping force required when an earthquake occurs.

The railroad vehicle vibration damping device of the present invention has a damper and either a semi-active damper or an actuator that are interposed in parallel between the body and bogie of the railroad vehicle, and the required damping force is the sum of the damping forces generated by the damper and either the semi-active damper or the actuator when expanding and contracting at a predetermined stroke speed expected when an earthquake occurs, and the damper generates a damping force greater than the required damping force when expanding and contracting at a predetermined stroke speed or faster. With this railroad vehicle vibration damping device configured in this way, the entire railroad vehicle vibration damping device generates the required damping force required when an earthquake occurs using only the damper when expanding and contracting at the predetermined stroke speed expected when an earthquake occurs, so the semi-active damper (actuator) does not have to share the required damping force when expanding and contracting at the predetermined stroke speed expected when an earthquake occurs, and the semi-active damper (actuator) does not have to be designed to withstand high pressure.

The damper in the railroad vehicle vibration control device may also include a cylinder, a rod inserted into the cylinder so as to be movable in the axial direction, a piston attached to the rod and inserted into the cylinder so as to be movable in the axial direction, dividing the inside of the cylinder into a rod side chamber and a piston side chamber, a tank, a first orifice provided in a first damping passage that communicates between the rod side chamber and the piston side chamber, a second orifice provided in a second damping passage that communicates between the piston side chamber and the tank, a compression side check valve provided in a compression side suction passage that communicates between the rod side chamber and the tank and that allows only the flow of liquid from the tank to the rod side chamber, an extension side check valve provided in an extension side suction passage that communicates between the piston side chamber and the tank and that allows only the flow of liquid from the tank to the piston side chamber, and a compression side check valve provided in a compression side suction passage that communicates between the piston side chamber and the tank and that allows only the flow of liquid from the tank to the piston side chamber.

With a railway vehicle vibration control device having a damper configured in this way, when the damper is contracting, the first orifice increases the pressure in the piston side chamber while the compression side check valve opens to make the pressure in the rod side chamber equal to the tank pressure. This means that the pressure-receiving area of the piston that is effective for generating a damping force when the damper is contracting is the cross-sectional area of the piston itself, and because a larger pressure-receiving area can be secured compared to a uniflow damper, less pressure is required in the piston side chamber to generate the same amount of damping force, and the pressure resistance of each component that makes up the damper can be reduced, reducing product costs.

Furthermore, in the railway vehicle vibration control device of this embodiment, the damper and either the semi-active damper or the actuator each have a cylinder, a rod inserted into the cylinder so as to be movable in the axial direction, and a piston attached to the rod and inserted into the cylinder so as to be movable in the axial direction, dividing the inside of the cylinder into a rod side chamber and a piston side chamber, and the inner diameter of the damper cylinder may be larger than the inner diameter of the semi-active damper or one of the actuator cylinders.

With a railway vehicle vibration control device configured in this way, the damper can generate a greater damping force than a semi-active damper (actuator), while suppressing high pressure inside the cylinder, so the requirements for the load-bearing capacity and sealing performance of each part of the damper are not as strict, and the product cost of the entire system can be reduced compared to conventional products that must adopt a pressure-resistant structure. Also, it is only necessary to increase the inner diameter of the cylinder of the damper alone, and there is no need to increase the inner diameter of the cylinder of the semi-active damper (actuator).

Furthermore, in the railway vehicle vibration control device, the damper and either the semi-active damper or the actuator each have a cylinder, a rod inserted into the cylinder so as to be movable in the axial direction, and a piston attached to the rod and inserted into the cylinder so as to be movable in the axial direction, dividing the inside of the cylinder into a rod side chamber and a piston side chamber, and the thickness of the damper cylinder may be made thicker than the thickness of the semi-active damper or one of the actuator cylinders.

With a railcar vibration damping device configured in this way, the strength of the cylinder in the damper can be improved and pressure inside the cylinder can be increased without increasing the outer diameter compared to a semi-active damper (actuator), and the railcar vibration damping device can be mounted on a railcar without compromising its mountability.

The railway vehicle vibration damping device may be such that the damper and either the semi-active damper or the actuator each have a cylinder, a rod inserted into the cylinder so as to be movable in the axial direction, and a piston attached to the rod and inserted into the cylinder so as to be movable in the axial direction, dividing the interior of the cylinder into a rod side chamber and a piston side chamber, and the strength of the material of the cylinder, rod and piston of the damper may be greater than the strength of the material of the cylinder, rod and piston of the semi-active damper or the actuator.

With a railroad vehicle vibration damping device configured in this way, the strength of the damper can be improved and pressure inside the cylinder can be increased without increasing the outer diameter compared to a semi-active damper (actuator), and the railroad vehicle vibration damping device can be easily installed on a railroad vehicle.

The railway vehicle vibration control device of the present invention is capable of generating the necessary damping force required when an earthquake occurs, but does not require the semi-active damper or actuator to be designed to withstand high pressures, reducing the product cost of the entire system.

FIG. 2 is a diagram showing a railway vehicle vibration damping device mounted on a railway vehicle. FIG. 2 is a hydraulic circuit diagram of a damper according to an embodiment. 5 is a diagram showing a damping force characteristic, which is a characteristic of the damping force with respect to the expansion/contraction speed of the damper in one embodiment. FIG. FIG. 2 is a hydraulic circuit diagram of a semi-active damper according to one embodiment. FIG. 2 is a hydraulic circuit diagram of an actuator damper according to one embodiment.

The present invention will be described below based on the embodiment shown in the drawings. In one embodiment, a railway vehicle vibration damping device S is used to damp vibrations in a railway vehicle car body B. As shown in FIG. 1, the railway vehicle vibration damping device S in this embodiment includes a damper D interposed between the car body B and the bogie T, and either a semi-active damper SD or an actuator A.

As shown in FIG. 1, the railroad vehicle vibration damping device S of this embodiment is used as a vibration damping device that suppresses left-right lateral vibrations of the car body B of the railroad vehicle to which it is to be installed in the vehicle travel direction. The damper D and one of the semi-active damper SD or actuator A are connected to a pin P that hangs down below the car body B, and are interposed in parallel between the car body B and the bogie T in a pair. The bogie T holds the wheels W rotatably, and a suspension spring CS is interposed between the car body B and the bogie T, and the car body B is elastically supported from below, allowing the car body B to move in the lateral direction relative to the bogie T. The railroad vehicle vibration damping device S is used to suppress lateral vibrations of the bogie T relative to the car body B in the vehicle travel direction, but may also be used to suppress vibrations of the bogie T swinging horizontally around the pin P relative to the car body B.

<Damper>
First, the damper D will be described. As shown in Fig. 2, the damper D includes a cylinder 1, a rod 2 inserted into the cylinder 1 so as to be movable in the axial direction, a piston 3 attached to the rod 2 and inserted into the cylinder 1 so as to be movable in the axial direction, dividing the inside of the cylinder 1 into a rod side chamber 4 and a piston side chamber 5, a tank 6, a first orifice 8 provided in a first damping passage 7 communicating between the rod side chamber 4 and the piston side chamber 5, a second orifice 10 provided in a second damping passage 9 communicating between the piston side chamber 5 and the tank 6, and a compression side check valve 12 provided in a compression side suction passage 11 communicating between the rod side chamber 4 and the tank 6, which allows only the flow of liquid from the tank 6 to the rod side chamber 4. an extension-side check valve 14 provided in an extension-side suction passage 13 communicating between the piston side chamber 5 and the tank 6, and allowing only the flow of liquid from the tank 6 to the piston side chamber 5; an extension-side relief valve 16 provided in an extension-side relief passage 15 communicating between the rod side chamber 4 and the tank 6, and opening when the pressure in the rod side chamber 4 reaches a valve opening pressure to allow the flow of liquid from the rod side chamber 4 to the tank 6; and a compression-side relief valve 18 provided in a compression-side relief passage 17 communicating between the piston side chamber 5 and the tank 6, and opening when the pressure in the piston side chamber 5 reaches a valve opening pressure to allow the flow of liquid from the piston side chamber 5 to the tank 6.

The cylinder 1 is cylindrical and is housed in an outer cylinder 19 that covers the outer circumference of the cylinder 1, forming an annular tank 6 between the cylinder 1 and the outer cylinder 19. The openings on the left end side of the cylinder 1 and the outer cylinder 19 in FIG. 2 are closed by an annular rod guide 20 that fits into both, and the openings on the right end side of the cylinder 1 and the outer cylinder 19 in FIG. 2 are closed by a bottom cap 21 that fits into both. In addition, a rod 2 that is inserted movably into the cylinder 1 is slidably inserted into the rod guide 20, and the axial movement of the rod 2 is guided by the rod guide 20. In addition, the rod 2 is inserted axially movably into the cylinder 1, and one end of the rod 2 protrudes outside the cylinder 1 through the rod guide 20, and the other end protrudes into the cylinder 1, and is connected to a piston 3 that is inserted axially movably into the cylinder 1.

The outer circumference of the rod 2, the gap between the rod guide 20 and the cylinder 1, the gap between the rod guide 20 and the outer tube 19, the gap between the cylinder 1 and the bottom cap 21, and the gap between the outer tube 19 and the bottom cap 21 are each sealed with sealing members (not shown). This keeps the inside of the cylinder 1 and the tank 6 sealed.

The inside of the cylinder 1 is divided into a rod side chamber 4 and a piston side chamber 5 by a piston 3 inserted into the cylinder 1. In this embodiment, the rod side chamber 4 and the piston side chamber 5 are filled with hydraulic oil as a liquid, and the tank 6 is filled with gas in addition to storing hydraulic oil. It is not necessary to compress and fill the tank 6 with gas to create a pressurized state. The liquid may be a liquid other than hydraulic oil.

In this way, the damper D of this embodiment is a so-called single-rod type damper in which the rod 2 protrudes from only one end side of the cylinder 1, so it is easier to ensure the stroke length compared to a double-rod type damper, and the overall length of the damper D is shorter, improving the mountability on a railway vehicle.

In addition, a bracket (not shown) is provided on the left end of the rod 2 in FIG. 2 and on the bottom cap 21 that closes the right end of the cylinder 1, so that the damper D can be interposed between the car body B and the bogie T of the railway vehicle.

As shown in FIG. 2, the first damping passage 7 is provided in the piston 3 and connects the rod side chamber 4 and the piston side chamber 5. The first orifice 8 is provided in the piston 3 and provides resistance to the flow of hydraulic oil passing through the first damping passage 7. In addition to the first damping passage 7 and the first orifice 8, the piston 3 in the damper D of this embodiment is provided with a third damping passage 22 that connects the rod side chamber 4 and the piston side chamber 5, a third orifice 23 provided midway through the third damping passage 22, and a third damping passage check valve 24 that is provided midway through the third damping passage 22 and allows only the flow of hydraulic oil from the piston side chamber 5 to the rod side chamber 4.

The second damping passage 9 connects the piston side chamber 5 to the tank 6. The second orifice 10 is installed in the bottom cap 21 and provides resistance to the flow of hydraulic oil passing through the second damping passage 9. The compression side suction passage 11 connects the rod side chamber 4 to the tank 6. The compression side check valve 12 is installed in the compression side suction passage 11 and allows only hydraulic oil to flow through the compression side suction passage 11 from the tank 6 to the rod side chamber 4, setting the compression side suction passage 11 as a one-way passage that allows only hydraulic oil to flow from the tank 6 to the rod side chamber 4.

The extension-side suction passage 13 connects the piston-side chamber 5 to the tank 6. The extension-side check valve 14 is provided in the extension-side suction passage 13 and allows hydraulic oil to flow only from the tank 6 to the piston-side chamber 5 through the extension-side suction passage 13, setting the extension-side suction passage 13 as a one-way passage that allows hydraulic oil to flow only from the tank 6 to the piston-side chamber 5.

The extension-side relief passage 15 connects the rod-side chamber 4 to the tank 6. The extension-side relief valve 16 is provided in the extension-side relief passage 15, and opens when the pressure in the rod-side chamber 4 reaches the valve opening pressure to allow hydraulic oil to flow from the rod-side chamber 4 to the tank 6, setting the extension-side relief passage 15 as a one-way passage that only allows hydraulic oil to flow from the rod-side chamber 4 to the tank 6.

Furthermore, the compression side relief passage 17 connects the piston side chamber 5 to the tank 6. The compression side relief valve 18 is provided in the compression side relief passage 17, and opens when the pressure in the piston side chamber 5 reaches the valve opening pressure to allow the flow of hydraulic oil from the piston side chamber 5 to the tank 6, setting the compression side relief passage 17 as a one-way passage that only allows the flow of hydraulic oil from the piston side chamber 5 to the tank 6.

The damper D of this embodiment is also provided with a small diameter orifice 25 that has a smaller diameter than the first orifice 8 that connects the rod side chamber 4 and the tank 6. The small diameter orifice 25 functions as an air vent that discharges air bubbles into the tank 6 when air bubbles are generated in the hydraulic oil in the cylinder 1. The small diameter orifice 25 may be omitted if not required.

When the damper D configured in this way performs an extension operation in which the rod 2 moves leftward in FIG. 2 relative to the cylinder 1, the piston 3 moves leftward in the cylinder 1, contracting the rod side chamber 4 and expanding the piston side chamber 5. If the stroke speed during the extension operation of the damper D is low and the pressure in the contracted rod side chamber 4 does not reach the opening pressure of the extension side relief valve 16, the hydraulic oil in the contracted rod side chamber 4 passes only through the first damping passage 7 and moves to the expanding piston side chamber 5, and the first orifice 8 provides resistance to the flow of hydraulic oil, so that the pressure in the rod side chamber 4 becomes higher than the pressure in the piston side chamber 5. In addition, during the extension operation of the damper D, the rod 2 exits from the cylinder 1, so that there is a shortage of hydraulic oil in the cylinder 1 by the volume of the rod 2 exiting from the cylinder 1. However, the extension side check valve 14 opens and the shortage of hydraulic oil is supplied from the tank 6 to the piston side chamber 5 via the extension side suction passage 13. In this way, until the damper D extends and the pressure in the rod side chamber 4 reaches the opening pressure of the extension relief valve 16, the first orifice 8 of the damper D generates a damping force that resists the flow of hydraulic oil passing through the first damping passage 7, suppressing the extension operation.

On the other hand, when the damper D extends and the pressure in the rod side chamber 4 reaches the opening pressure of the extension side relief valve 16, the extension side relief valve 16 opens to suppress the increase in pressure in the rod side chamber 4, so that the damper D generates a damping force that impedes the extension operation mainly by the resistance that the extension side relief valve 16 provides to the hydraulic oil passing through.

2 relative to the cylinder 1, the piston 3 moves to the right in the cylinder 1, contracting the piston side chamber 5 and expanding the rod side chamber 4. If the stroke speed during the contraction of the damper D is low and the pressure in the contracted piston side chamber 5 does not reach the opening pressure of the compression side relief valve 18, the hydraulic oil in the contracted piston side chamber 5 passes through the first damping passage 7, the second damping passage 9, and the third damping passage 22 and moves to the expanding rod side chamber 4 and the tank 6, and the first orifice 8, the second orifice 10, and the third orifice 23 provide resistance to the flow of the hydraulic oil, so that the pressure in the piston side chamber 5 becomes higher than the pressure in the rod side chamber 4. During the contraction of the damper D, the rod 2 enters the cylinder 1, so that the hydraulic oil in the cylinder 1 is excessive by the volume of the rod 2 entering the cylinder 1. However, the excess hydraulic oil is discharged from the piston side chamber 5 to the tank 6 through the second damping passage 9 and the second orifice 10. The rod side chamber 4 is connected to the tank 6 through the compression side suction passage 11. During the contraction of the damper D, the compression side check valve 12 opens and hydraulic oil is supplied from the tank 6 to the expanding rod side chamber 4, so that the pressure in the rod side chamber 4 becomes the tank pressure. In this way, until the damper D contracts and the pressure in the piston side chamber 5 reaches the opening pressure of the compression side relief valve 18, the damper D generates a damping force that suppresses the contraction by providing resistance to the flow of hydraulic oil passing through the second damping passage 9 and the third damping passage 22 through the second orifice 10 and the third orifice 23.

On the other hand, when the damper D contracts and the pressure in the piston side chamber 5 reaches the opening pressure of the compression side relief valve 18, the compression side relief valve 18 opens to suppress the pressure rise in the piston side chamber 5, so that the damper D generates a damping force that prevents the contraction operation mainly by the resistance that the compression side relief valve 18 provides to the hydraulic oil passing through.

When the damper D is extended, the third damping passage check valve 24 closes to prevent hydraulic oil from passing through the third damping passage 22. Therefore, the third orifice 23 functions as a damping force source only when the damper D is contracted, and does not affect the damping force when the damper D is extended. When the damper D is contracted, the second orifice 10 and the third orifice 23 can increase the pressure in the piston side chamber 5, and the pressure in the rod side chamber 4 can be made the tank pressure by opening the compression side check valve 12. Therefore, when the damper D is contracted, if the tank pressure is considered to be 0, the entire surface of the piston 3 facing the piston side chamber 5 can be used as a pressure receiving area to generate a damping force equal to the pressure in the piston side chamber 5 multiplied by the cross-sectional area of the piston 3.

Thus, in the damper D of this embodiment, compared to a uniflow-type damper in which hydraulic oil is discharged from the rod side chamber to the tank whether it is extended or contracted, and the hydraulic oil circulates through the tank, piston side chamber, and rod side chamber in sequence, when the damper D is contracted, the pressure in the rod side chamber 4 can be made the tank pressure while increasing the pressure in the piston side chamber 5. Therefore, in the damper D, the pressure-receiving area of the piston 3 that is effective for generating a damping force when it is contracted is the cross-sectional area of the piston 3 itself, and because a large pressure-receiving area can be ensured, only a small pressure is required in the piston side chamber 5 to generate the same magnitude of damping force. Incidentally, the third damping passage 22, the third orifice 23 and the third damping passage check valve 24 may be eliminated. In that case, the operation of the damper D during extension is similar to that described above. However, when the damper D is contracting and the compression side relief valve 18 is not opened, the hydraulic oil in the piston side chamber 5 being contracted moves through the first damping passage 7 and the second damping passage 9 to the rod side chamber 4 and the tank 6, and the damper D generates a damping force through the first orifice 8 and the second orifice 10. When the damper D is contracting and the compression side relief valve 18 is opened, the hydraulic oil in the piston side chamber 5 being contracted moves through the first damping passage 7, the second damping passage 9 and the compression side relief passage 17 to the rod side chamber 4 and the tank 6, and the damper D generates a damping force mainly through the compression side relief valve 18. In this way, even if the third damping passage 22, the third orifice 23, and the third damping passage check valve 24 are eliminated, the pressure in the rod side chamber 4 can be set to tank pressure while increasing the pressure in the piston side chamber 5 during the contraction operation of the damper D. Therefore, the pressure-receiving area of the piston 3 that is effective for generating a damping force during the contraction operation of the damper D is the cross-sectional area of the piston 3 itself, and since a larger pressure-receiving area can be secured compared to a uniflow type damper, the pressure in the piston side chamber 5 required to generate the same magnitude of damping force is smaller.

In addition, the damper D of this embodiment is provided with a third orifice 23 and a third damping passage check valve 24. The damper D is required to generate the necessary damping force, and there is a limit to how large the diameter of the first orifice 8 can be made. Since the amount of hydraulic oil moving from the piston side chamber 5 to the rod side chamber 4 during the contraction operation of the damper D is greater than the amount of hydraulic oil moving from the rod side chamber 4 to the piston side chamber 5 during the extension operation of the damper D, if the third orifice 23 is eliminated and only the first orifice 8 is used, the difference between the pressure in the piston side chamber 5 and the pressure in the rod side chamber 4 during the contraction operation of the damper D may become large. However, by providing the third orifice 23 that is effective only during the contraction operation of the damper D, the difference between the pressure in the piston side chamber 5 and the pressure in the rod side chamber 4 during the contraction operation of the damper D can be appropriately adjusted, so that the damping force characteristics during the contraction operation of the damper D can be easily tuned. Furthermore, in the damper D of this embodiment that includes the third orifice 23 and the third damping passage check valve 24, a check valve (not shown) that allows only the flow of hydraulic oil from the rod side chamber 4 to the piston side chamber 5 may be provided in the first damping passage 7 so that the first orifice 8 in the first damping passage 7 generates a damping force only during the extension operation of the damper D, without affecting the damping force during the contraction operation of the damper D. In this case, the first orifice 8 generates a damping force during the extension operation of the damper D, and the second orifice 10 and the third orifice 23 generate a damping force during the contraction operation of the damper D. Since the size of the aperture of the third orifice 23 can be determined regardless of the aperture of the first orifice 8, the damping force characteristics during the contraction operation of the damper D can be easily tuned. Even in this case, the third orifice 23 can create a difference between the pressure in the piston side chamber 5 and the pressure in the rod side chamber 4 when the damper D is contracting, and since a larger pressure-receiving area can be secured compared to a uniflow type damper, less pressure is required in the piston side chamber 5 to generate the same amount of damping force.

In addition, the damper D of this embodiment is provided with an extension side relief valve 16 that is provided in the extension side relief passage 15 that communicates between the rod side chamber 4 and the tank 6 and opens when the pressure in the rod side chamber 4 reaches the valve opening pressure to allow the flow of liquid from the rod side chamber 4 to the tank 6, and a compression side relief valve 18 that is provided in the compression side relief passage 17 that communicates between the piston side chamber 5 and the tank 6 and opens when the pressure in the piston side chamber 5 reaches the valve opening pressure to allow the flow of liquid from the piston side chamber 5 to the tank 6. Therefore, even if the stroke speed during the contraction operation of the damper D becomes faster than the predetermined stroke speed, the extension side relief valve 16 or the compression side relief valve 18 opens, and the pressure in the compression side chamber in the cylinder 1 is prevented from becoming a pressure that the seal members and each member cannot tolerate.

The railway vehicle vibration control device S of this embodiment is equipped with a damper D and a semi-active damper SD that are interposed in parallel between the car body B and the bogie T, and the damper D generates a damping force greater than the required damping force when it expands and contracts at a predetermined stroke speed or faster, with the required damping force being the sum of the damping forces generated by the damper D and the semi-active damper SD required when expanding and contracting at a predetermined stroke speed expected in the event of an earthquake.

Specifically, the predetermined stroke speed is set to the stroke speed of the damper D at which the railcar may derail, and in this embodiment, it is set to 0.6 m/sec, and the required damping force is 80 kN. Therefore, as shown in FIG. 3, the damper D is set to exert a damping force of at least 80 kN when the stroke speed reaches the predetermined stroke speed of 0.6 m/sec.

In this embodiment, the stroke speed when the damper D extends and the pressure in the rod side chamber 4 reaches the opening pressure of the extension relief valve 16 is called the valve opening speed, and the valve opening speed is set to about 0.5 m/sec, which is slightly lower than the predetermined stroke speed. Therefore, as shown in Figure 3, the damping force characteristic, which is the characteristic of the damping force against the stroke speed during the extension operation of the damper D, is proportional to the square of the stroke speed specific to the orifice, because the first orifice 8 generates a damping force that hinders the extension operation when the stroke speed is lower than the valve opening speed.

On the other hand, when the stroke speed during the extension operation of the damper D becomes equal to or exceeds the opening speed, the extension relief valve 16 opens, and the pressure in the rod side chamber 4 becomes equal to the opening pressure of the extension relief valve 16 plus a pressure override according to the amount of hydraulic oil passing through the extension relief valve 16. The pressure override increases in proportion to the amount of hydraulic oil passing through the extension relief valve 16, but is smaller than the pressure loss due to the orifice. Therefore, as shown in Figure 3, the damping force characteristic of the damper D when the stroke speed during the extension operation becomes equal to or exceeds the opening speed is such that the damping force increases with increasing stroke speed, but the damping coefficient is smaller than the damping force characteristic when the stroke speed is less than the opening speed, and the increase in damping force reaches a plateau.

In this embodiment, when the stroke speed of the damper D during extension operation reaches a predetermined stroke speed of 0.6 m/sec, the extension side relief valve 16 is set so that the pressure in the rod side chamber 4 becomes the pressure required to make the damping force generated by the damper D 80 kN, and the damper D generates a damping force equal to the required damping force when it extends at the predetermined stroke speed.

In this embodiment, the stroke speed when the damper D contracts and the pressure in the piston side chamber 5 reaches the opening pressure of the compression side relief valve 18 is set as the valve opening speed, and the valve opening speed during contraction is set to about 0.5 m/sec, which is the same as the valve opening speed during extension. Therefore, as shown in FIG. 3, when the stroke speed is lower than the valve opening speed, the first orifice 8, the second orifice 10, and the third orifice 23 generate a damping force that hinders the contraction, so the damping force is proportional to the square of the stroke speed specific to the orifice. The damper D in the vibration damping device S for railway vehicles is used for the purpose of suppressing the lateral vibration of the bogie T relative to the car body B, and the same damping force is required whether the bogie T moves to the left or right relative to the car body B in the vehicle travel direction, so the damping force characteristic during contraction of the damper D is set to be almost the same as the damping force characteristic during extension of the damper D.

On the other hand, when the stroke speed during the contraction operation of the damper D becomes equal to or exceeds the opening speed, the compression side relief valve 18 opens, and the pressure in the piston side chamber 5 becomes equal to the opening pressure of the compression side relief valve 18 plus a pressure override according to the amount of hydraulic oil passing through the compression side relief valve 18. The pressure override increases in proportion to the amount of hydraulic oil passing through the compression side relief valve 18, but is smaller than the pressure loss due to the orifice. Therefore, as shown in Figure 3, the damping force characteristics of the damper D when the stroke speed during the contraction operation becomes equal to or exceeds the opening speed are such that the damping force increases with increasing stroke speed, but compared to the damping force characteristics when the stroke speed is less than the opening speed, the damping coefficient is small and the increase in damping force plateaus.

In this embodiment, when the stroke speed of the damper D during contraction reaches a predetermined stroke speed of 0.6 m/sec, the compression side relief valve 18 is set so that the pressure in the piston side chamber 5 becomes the pressure required to make the damping force generated by the damper D 80 kN, and the damper D generates a damping force equal to the required damping force when contracting at the predetermined stroke speed.

The opening speed of the damper D during the expansion/contraction operation of the expansion-side relief valve 16 and the compression-side relief valve 18 is set lower than the predetermined stroke speed, but the opening speed may be set equal to or higher than the predetermined stroke speed. In that case, the required damping force to be generated when the damper D expands and contracts at the predetermined stroke speed can be generated by the first orifice 8 during the expansion operation, and by the first orifice 8, the second orifice 10, and the third orifice 23 during the contraction operation. If the third orifice 23 is eliminated, the required damping force to be generated when the damper D contracts at the predetermined stroke speed can be generated by the first orifice 8 and the second orifice 10.

In addition, the first orifice 8, the second orifice 10, and the third orifice 23 cause the damper D to exert a damping force of, for example, 15 kN or less when the extension/contraction speed of the damper D is in a normal speed range that is lower than the normal maximum speed, which is lower than the specified stroke speed. Specifically, the normal maximum speed is set to the maximum stroke speed of the damper D that would be reached when a railway vehicle runs in a state where no earthquake occurs, and is set to 0.2 m/sec in this embodiment. Therefore, the normal speed range of the stroke speed when the damper D is extended/contracted is set to a range from 0 to 0.2 m/sec, and the first orifice 8, the second orifice 10, and the third orifice 23 are set so that the damping force generated by the damper D when it is extended/contracted at a stroke speed within the normal speed range is 15 kN or less.

Therefore, the damping force characteristic (damping force characteristic) with respect to the extension/contraction speed of damper D is a square characteristic specific to an orifice, as shown in Figure 3, and is less than 15 kN when the stroke speed of damper D is in the range of less than 0.2 m/sec, which is the normal speed range, and is 80 kN or more when the stroke speed of damper D is equal to or greater than the specified stroke speed of 0.6 m/sec.

In this embodiment, the specified stroke speed is set to 0.6 m/sec, and the normal maximum speed is set to 0.2 m/sec, but both may be set to other values. In addition, the required damping force may be set to a value that can prevent the vehicle from running off the track, depending on the specifications of the railcar.

<Semi-active damper>
As shown in FIG. 4 , the semi-active damper SD includes a cylinder 31 connected to a car body B of a railway vehicle, a rod 32 inserted into the cylinder 31 so as to be movable in the axial direction and one end of which is connected to a bogie T, a piston 33 inserted into the cylinder 31 so as to be movable in the axial direction, dividing the inside of the cylinder 31 into a rod side chamber 34 and a piston side chamber 35, and connected to the other end of the rod 32, a tank 36, a first passage 37 communicating between the rod side chamber 34 and the piston side chamber 35, and a piston 33 provided in the first passage 37. the first passage 37; a second passage 39 connecting the piston side chamber 35 and the tank 36; a second on-off valve 40 provided in the second passage 39 for opening and closing the second passage 39; a discharge passage 41 connecting the rod side chamber 34 and the tank 36; a variable relief valve 42 provided in the discharge passage 41; a flow straightening passage 43 allowing only a flow from the piston side chamber 35 to the rod side chamber 34; and a suction passage 44 allowing only a flow from the tank 36 to the piston side chamber 35.

Each part of the semi-active damper SD will be described in detail below. The cylinder 31 is cylindrical and is housed in an outer cylinder 45 that covers the outer circumference of the cylinder 1, and forms an annular tank 36 between the cylinder 31 and the outer cylinder 45. The openings on the left end side of the cylinder 31 and the outer cylinder 45 in FIG. 4 are closed by an annular rod guide 46 that fits into both, and the openings on the right end side of the cylinder 31 and the outer cylinder 45 in FIG. 4 are closed by a bottom cap 47 that fits into both. In addition, the rod 32 that is inserted movably into the cylinder 31 is slidably inserted into the rod guide 46, and the axial movement of the rod 32 is guided by the rod guide 46. In addition, the rod 32 is inserted axially movably into the cylinder 31, and one end of the rod 32 protrudes outside the cylinder 31 through the rod guide 46, and the other end of the rod 32 protrudes into the cylinder 31, and is connected to a piston 33 that is inserted axially movably into the cylinder 31.

The outer periphery of the rod 32, the gap between the rod guide 46 and the cylinder 31, the gap between the rod guide 46 and the outer tube 45, the gap between the cylinder 31 and the bottom cap 47, and the gap between the outer tube 45 and the bottom cap 47 are each sealed with sealing members (not shown). This keeps the inside of the cylinder 31 and the tank 36 sealed.

The inside of the cylinder 31 is divided into a rod side chamber 34 and a piston side chamber 35 by a piston 33 inserted into the cylinder 31. In this embodiment, the rod side chamber 34 and the piston side chamber 35 are filled with hydraulic oil as a liquid, and the tank 36 is filled with gas in addition to storing hydraulic oil. It is not necessary to compress and fill the gas in the tank 36 to create a pressurized state. The liquid may be a liquid other than hydraulic oil.

In this way, the semi-active damper SD of this embodiment is a so-called single-rod type damper in which the rod 32 protrudes from only one end side of the cylinder 31, so it is easier to ensure the stroke length compared to a double-rod type damper, and the overall length of the semi-active damper SD is shorter, improving the mountability on railway vehicles.

In addition, in the case of this semi-active damper SD, the cross-sectional area of the rod 32 is half that of the piston 33, so that the pressure-receiving area on the rod side chamber 34 side of the piston 33 is half that of the piston side chamber 35 side. Therefore, the flow rate discharged from inside the cylinder 31 to the tank 36 through the discharge passage 41 is equal during the extension operation and the contraction operation of the semi-active damper SD.

In addition, a bracket (not shown) is provided on the left end of the rod 32 in FIG. 4 and on the bottom cap 47 that closes the right end of the cylinder 31, so that this semi-active damper SD can be interposed between the car body B and the bogie T of the railway vehicle.

The first passage 37 connects the rod side chamber 34 and the piston side chamber 35, and is provided with a first on-off valve 38 in the middle. This first passage 37 connects the rod side chamber 34 and the piston side chamber 35 outside the cylinder 31, but may be provided in the piston 33.

The first on-off valve 38 is an electromagnetic on-off valve, and has a communication position in which the first passage 37 is opened to connect the rod side chamber 34 and the piston side chamber 35, and a blocking position in which the first passage 37 is blocked to cut off communication between the rod side chamber 34 and the piston side chamber 35. When energized, the first passage 37 is opened to connect the rod side chamber 34 and the piston side chamber 35.

The second passage 39 also connects the piston side chamber 35 to the tank 36, and is provided with a second on-off valve 40 in the middle. The second on-off valve 40 is an electromagnetic on-off valve, and has a connection position where the second passage 39 is opened to connect the piston side chamber 35 to the tank 36, and a disconnection position where the second passage 39 is blocked to cut off communication between the piston side chamber 35 and the tank 36. When energized, the second passage 39 is opened to connect the piston side chamber 35 to the tank 36.

The discharge passage 41 communicates between the rod side chamber 34 and the tank 36, and is provided with a variable relief valve 42 that can change the valve opening pressure. In this embodiment, the variable relief valve 42 is a proportional electromagnetic relief valve that opens when the pressure in the rod side chamber 34 reaches or exceeds the valve opening pressure, allowing the flow of hydraulic oil from the rod side chamber 34 to the tank 36. The variable relief valve 42 can adjust the valve opening pressure according to the amount of current supplied, and is configured to minimize the valve opening pressure when the amount of current is maximized, and maximize the valve opening pressure when no current is supplied.

The straightening passage 43 communicates between the piston side chamber 35 and the rod side chamber 34, and is provided with a check valve 43a that only allows hydraulic oil to flow from the piston side chamber 35 to the rod side chamber 34, and is set as a one-way passage that only allows hydraulic oil to flow from the piston side chamber 35 to the rod side chamber 34. The suction passage 44 communicates between the tank 36 and the piston side chamber 35, and is provided with a check valve 44a that only allows hydraulic oil to flow from the tank 36 to the piston side chamber 35, and is set as a one-way passage that only allows hydraulic oil to flow from the tank 36 to the piston side chamber 35.

In the semi-active damper SD of this embodiment, when the first opening/closing valve 38 and the second opening/closing valve 40 are in the shutoff position and the semi-active damper SD is extended by an external force, hydraulic oil is pushed out from the rod side chamber 34, which is contracted by the piston 33 moving in the cylinder 31, to the tank 36 through the discharge passage 41. In addition, hydraulic oil is supplied from the tank 36 through the suction passage 44 to the piston side chamber 35, which is expanded by the movement of the piston 33 in the cylinder 31. During this extension operation, the semi-active damper SD exerts a damping force that suppresses extension by applying resistance to the flow of hydraulic oil with the variable relief valve 42, and the damping force can be adjusted by adjusting the valve opening pressure of the variable relief valve 42.

On the other hand, in the semi-active damper SD of this embodiment, when the first opening/closing valve 38 and the second opening/closing valve 40 are in the shutoff position, when the semi-active damper SD is contracted by an external force, hydraulic oil moves from the piston side chamber 35, which is contracted by the piston 33 moving in the cylinder 31, to the rod side chamber 34, which is expanded through the flow straightening passage 43. Also, when the semi-active damper SD is contracting, the rod 32 enters the cylinder 31, so that the hydraulic oil in the cylinder 31 by the volume of the rod 32 entering the cylinder 31 becomes excessive and is discharged to the tank 36 through the discharge passage 41. During this contraction operation, the semi-active damper SD exerts a damping force that suppresses the contraction by applying resistance to the flow of hydraulic oil through the variable relief valve 42, just as during the extension operation, so that the damping force can be adjusted by adjusting the valve opening pressure of the variable relief valve 42.

In this case, the flow rate of hydraulic oil passing through the discharge passage 41 is equal to the cross-sectional area of the rod 32 multiplied by the amount of movement of the piston 33. Here, the cross-sectional area of the rod 32 is set to half the cross-sectional area of the piston 33, so if the amount of movement of the piston 33 is the same whether the semi-active damper SD is extended or contracted, the flow rate of hydraulic oil passing through the discharge passage 41 will be the same. Therefore, the semi-active damper SD can exert the same damping force if the movement speed of the piston 33 is the same on both sides of the extension and contraction.

The first and second opening/closing valves 38 and 40 are in the shutoff position when not energized, and in the event of a failure in which power cannot be supplied to the first and second opening/closing valves 38 and 40, the semi-active damper SD of this embodiment always exerts a damping force against expansion and contraction as described above, and therefore functions as a passive damper. In this way, when power cannot be supplied, the variable relief valve 42 maximizes the valve opening pressure, so that the semi-active damper SD maximizes the damping force in the event of a failure in which power cannot be supplied.

In addition, in the semi-active damper SD of this embodiment, when the first opening/closing valve 38 is in the communication position and the second opening/closing valve 40 is in the blocking position, the rod side chamber 34 and the piston side chamber 35 are in communication with each other through the first passage 37, but the communication between the piston side chamber 35 and the tank 36 through the second passage 39 is cut off. In this state, when the semi-active damper SD is contracted by an external force, the hydraulic oil corresponding to the volume of the rod 32 entering the cylinder 31 is discharged from the cylinder 31 to the exhaust passage 41, and a damping force against the contraction is exerted in the same manner as described above. On the other hand, in this state, when the semi-active damper SD extends, the hydraulic oil moves from the contracting rod side chamber 34 to the expanding piston side chamber 35 through the first passage 37, and the hydraulic oil corresponding to the volume of the rod 32 exiting the cylinder 31 is supplied from the tank 36 to the cylinder 31 through the suction passage 44. Therefore, in this case, the hydraulic oil does not flow to the exhaust passage 41, so that the semi-active damper SD does not exert a damping force.

Furthermore, in the semi-active damper SD of this embodiment, when the first on-off valve 38 is in the shutoff position and the second on-off valve 40 is in the communication position, the communication between the rod side chamber 34 and the piston side chamber 35 via the first passage 37 is cut off, but the piston side chamber 35 and the tank 36 are communicated via the second passage 39. When the semi-active damper SD is extended in this state by an external force, hydraulic oil is discharged from the rod side chamber 34 to the discharge passage 41 as the rod side chamber 34 contracts, and a damping force against the extension is exerted as described above. On the other hand, when the semi-active damper SD contracts in this state, hydraulic oil moves from the contracting piston side chamber 35 to the expanding rod side chamber 34 via the straightening passage 43, and the hydraulic oil of the volume of the rod 32 entering the cylinder 31 is discharged from the piston side chamber 35 to the tank 36 via the second passage 39. Therefore, in this case, the hydraulic oil does not flow to the exhaust passage 41, so the semi-active damper SD does not exert a damping force. In this way, the semi-active damper SD can function as a one-sided damper that exerts a damping force by selecting either extension or contraction.

As described above, when the first opening/closing valve 38 and the second opening/closing valve 40 are in the shutoff position, the semi-active damper SD functions as a passive damper, and the damping force can be adjusted by adjusting the amount of current supplied to the variable relief valve 42 to adjust the valve opening pressure. In addition, when the first opening/closing valve 38 is in the open position and the second opening/closing valve 40 is in the shutoff position, and when the first opening/closing valve 38 is in the shutoff position and the second opening/closing valve 40 is in the open position, as described above, the semi-active damper SD is in a mode in which it exerts a damping force only for either extension or contraction, and the damping force can also be adjusted. Therefore, for example, if this mode is selected, when the direction in which the damping force is exerted is a direction in which the car body B is vibrated due to the vibration of the bogie T of the railway vehicle, the semi-active damper SD does not exert a damping force in such a direction, and can generate a damping force of an optimal magnitude for suppressing the vibration of the car body B only when it is effective for suppressing the vibration of the car body B. Therefore, this semi-active damper SD can easily achieve semi-active control based on Karnop's skyhook theory, and can therefore function as a skyhook semi-active damper.

The semi-active damper SD thus configured is designed to be used within the normal speed range described above, and the strength of each part of the semi-active damper SD is set so that it can generate the normal maximum damping force required to suppress vibration of the vehicle body B when it expands and contracts within the normal speed range. The normal maximum damping force is set to 15 kN, for example, and the strength of each part of the semi-active damper SD is set so that it can withstand the maximum load with a safety factor added so that it will not be destroyed even if 15 kN is generated. The semi-active damper SD may also be equipped with a relief passage that is parallel to the exhaust passage 41 and connects the rod side chamber 34 to the tank 36, and a relief valve that is provided in the relief passage and limits the pressure in the rod side chamber 34 so that the pressure in the rod side chamber 34 does not exceed the pressure when the maximum load described above is exerted.

<Actuator>
5, the actuator A includes a pump 50 capable of supplying hydraulic oil to the rod side chamber 34 in addition to the configuration of the semi-active damper SD. Specifically, the actuator A includes a supply passage 51 communicating between the tank 36 and the rod side chamber 34, a pump 50 provided in the supply passage 51 for sucking up hydraulic oil from the tank 36 and discharging it to the rod side chamber 34, and a check valve 52 provided on the discharge side of the pump 50 in the supply passage 51 for blocking the flow of hydraulic oil from the rod side chamber 34 to the tank 36. In this specification, both the force actively exerted by the actuator A and the force exerted in a direction hindering the expansion and contraction when the actuator A is expanded or contracted by an external force are forces used to suppress vibration of the car body B of the railway vehicle, and therefore these forces will be referred to as damping force without distinction.

The pump 50 is driven by a motor 53 controlled by a controller (not shown) and is a pump that discharges hydraulic oil in only one direction. The pump 50 is installed in the supply passage 51 with its suction port facing the tank 36 and its discharge port facing the rod side chamber 34. When driven by the motor 53, the pump 50 draws hydraulic oil from the tank 36 and supplies the hydraulic oil to the rod side chamber 34.

As mentioned above, the pump 50 discharges hydraulic oil in only one direction and does not switch rotation direction, so there is no problem with the discharge volume changing when the rotation direction is switched, and an inexpensive gear pump or the like can be used.

Furthermore, since the rotation direction of the pump 50 is always the same, the motor 53, which is the driving source that drives the pump 50, does not require high responsiveness to rotation switching, so a cheaper motor 53 can be used. The check valve 52 is provided to prevent the backflow of hydraulic oil to the pump 50 when the actuator A is forcibly expanded or contracted by an external force.

Next, when the actuator A configured as described above is to exert a desired damping force in the extension direction, the motor 53 is rotated to supply hydraulic oil from the pump 50 into the cylinder 31, while the first opening/closing valve 38 is set to the communication position and the second opening/closing valve 40 is set to the cut-off position. Then, the rod side chamber 34 and the piston side chamber 35 are placed in communication with each other via the first passage 37, hydraulic oil is supplied to both from the pump 50, the piston 33 is pushed to the left in FIG. 5, and the actuator A exerts a damping force in the extension direction. When the pressure in the rod side chamber 34 and the piston side chamber 35 exceeds the opening pressure of the variable relief valve 42, the variable relief valve 42 opens and the hydraulic oil is discharged to the tank 36 through the discharge passage 41. Therefore, the pressure in the rod side chamber 34 and the piston side chamber 35 is controlled to be equal to the opening pressure of the variable relief valve 42, which is determined by the amount of current applied to the variable relief valve 42. The actuator A exerts a damping force in the extension direction equal to the difference in pressure-receiving area between the piston side chamber 35 side and the rod side chamber 34 side of the piston 33 multiplied by the pressure in the rod side chamber 34 and the piston side chamber 35 controlled by the variable relief valve 42.

On the other hand, when actuator A is to exert a desired damping force in the contraction direction, the motor 53 is rotated to supply hydraulic oil from the pump 50 into the rod side chamber 34, while the first opening/closing valve 38 is set to the blocking position and the second opening/closing valve 40 is set to the communicating position. Then, the piston side chamber 35 and the tank 36 are placed in communication via the second passage 39, and hydraulic oil is supplied from the pump 50 to the rod side chamber 34, so that the piston 33 is pushed to the right in FIG. 5 and actuator A exerts a damping force in the contraction direction. Then, as described above, by adjusting the amount of current applied to the variable relief valve 42, actuator A exerts a damping force in the contraction direction multiplied by the pressure-receiving area on the rod side chamber 34 side of the piston 33 and the pressure in the rod side chamber 34 controlled by the variable relief valve 42.

In addition, when the rod 32 moves to the left in FIG. 5 due to an external force with the first opening/closing valve 38 open and the second opening/closing valve 40 closed, the actuator A does not exert a force in the direction that prevents the movement of the rod 32, that is, in the contraction direction, regardless of whether the pump 50 is driven or not. In this case, when the pump 50 is driven, the discharge flow rate of the pump 50 cannot keep up with the volume change in the cylinder 31 that decreases when the rod 32 retreats from the cylinder 31, and in that case, hydraulic oil is supplied from the tank 36 to the cylinder 31 through the suction passage 44. In addition, when the pump 50 is not driven, hydraulic oil is supplied from the tank 36 to the cylinder 31 through the suction passage 44 for the volume of the rod 32 retreating from the cylinder 31. In either case, the pressure in the cylinder 31 is the tank pressure, so the actuator A does not exert a damping force in the direction that prevents the movement of the rod 32, that is, in the contraction direction.

When the rod 32 moves to the right in FIG. 5 due to an external force with the first on-off valve 38 open and the second on-off valve 40 closed, the hydraulic oil pushed out of the cylinder 31 by the rod 32 entering the cylinder 31 is returned to the tank 36 through the discharge passage 41, regardless of whether the pump 50 is driven or not. In this case, the pressure inside the cylinder 31 is controlled to the desired pressure by the variable relief valve 42, so that the actuator A can exert a force in the direction that prevents the movement of the rod 32, that is, in the extension direction.

On the other hand, when the rod 32 moves to the right in FIG. 5 due to an external force with the first on-off valve 38 closed and the second on-off valve 40 open, the actuator A does not exert a force in a direction that prevents the movement of the rod 32, that is, in the extension direction, regardless of whether the pump 50 is driven or not. In this case, when the pump 50 is driven, the discharge flow rate of the pump 50 cannot keep up with the volume change in the rod side chamber 34 that increases when the rod 32 enters the cylinder 31, but in that case, hydraulic oil is supplied from the piston side chamber 35 to the rod side chamber 34 through the straightening passage 43. In addition, because the rod 32 enters the cylinder 31, the hydraulic oil in the cylinder 31 becomes excessive by the volume of the rod 32 entering the cylinder 31, but the excess hydraulic oil is discharged from the piston side chamber 35 on the compression side to the tank 36 through the second passage 39. In this case, even when the pump 50 is not in operation, hydraulic oil is supplied from the piston side chamber 35 through the straightening passage 43 to the volume of the rod side chamber 34 that increases when the rod 32 enters the cylinder 31, just as when the pump 50 is in operation. Excess hydraulic oil in the cylinder 31 caused by the rod 32 entering the cylinder 31 is discharged from the compressed piston side chamber 35 through the second passage 39 to the tank 36. In either case, the pressure in the cylinder 31 is the tank pressure, so the actuator A does not exert a damping force in the direction that prevents the movement of the rod 32, that is, in the extension direction.

When the rod 32 moves to the left in FIG. 6 due to an external force with the first on-off valve 38 closed and the second on-off valve 40 open, the hydraulic oil pushed out of the rod side chamber 34 is returned to the tank 36 through the discharge passage 41 regardless of whether the pump 50 is driven or not. In this case, the pressure in the rod side chamber 34 is controlled to the desired pressure by the variable relief valve 42, so that the actuator A can exert a force in the direction that prevents the movement of the rod 32, that is, in the contraction direction.

In other words, when the first on-off valve 38 is opened and the second on-off valve 40 is closed, or when the first on-off valve 38 is closed and the second on-off valve 40 is opened, the actuator A exerts a damping force in only one of the directions of extension or contraction in response to vibration input from an external force, regardless of the driving state of the pump 50.

Therefore, in the present embodiment, when the direction in which the force is exerted is a direction in which the vehicle body B is vibrated due to the vibration of the bogie T of the railway vehicle, the actuator A can function as a one-sided damper so as not to exert force in such a direction. Therefore, this actuator A can easily realize semi-active control based on Karnop's Skyhook theory, and can also function as a semi-active damper.

And, as can be understood from the explanation of the semi-active damper SD, even with this actuator A, it can function as a damper only by opening and closing the first opening and closing valve 38 and the second opening and closing valve 40. In other words, even in a situation where the pump 50 is driven by the motor 53, when the actuator A is forcibly expanded or contracted by an external force, it can function as a skyhook semi-active damper or a passive damper, and the damping force can be adjusted by adjusting the valve opening pressure of the variable relief valve 42. In this way, the actuator A not only functions as an actuator, but can also function as a damper only by opening and closing the first opening and closing valve 38 and the second opening and closing valve 40, regardless of the driving state of the motor 53. And, the direction in which the actuator A should exert a damping force is controlled only by opening and closing the first opening and closing valve 38 and the second opening and closing valve 40, and when the direction in which the actuator should exert a damping force is the same as the direction in which the actuator should exert a damping force as a semi-active damper, the opening and closing states of the first opening and closing valve 38 and the second opening and closing valve 40 coincide. Therefore, in actuator A, the state of the actuator and the skyhook semi-active damper can be switched without stopping and driving the pump 50 or performing the troublesome and abrupt switching operation of the first opening/closing valve 38 and the second opening/closing valve 40. Therefore, actuator A is a system with high responsiveness and reliability.

<Dimensions and strength of dampers and semi-active dampers (actuators)>
As described above, the damper D generates a damping force greater than the required damping force when expanding and contracting at a predetermined stroke speed or faster, with the required damping force being the sum of the damping forces generated by the damper D and the semi-active damper SD required when expanding and contracting at a predetermined stroke speed assumed when an earthquake occurs. In other words, the railway vehicle vibration control device S includes the damper D and the semi-active damper SD (actuator A) that are interposed in parallel between the car body B and the bogie T, and when expanding and contracting at a predetermined stroke speed assumed when an earthquake occurs, the entire railway vehicle vibration control device S can generate the required damping force required when an earthquake occurs using only the damper D. Therefore, the semi-active damper SD (actuator A) does not have to share the required damping force when expanding and contracting at a predetermined stroke speed assumed when an earthquake occurs.

For this reason, in the railway vehicle vibration control device S of this embodiment, the damper D is required to generate a greater required damping force to respond to earthquakes than the semi-active damper SD (actuator A), but the semi-active damper SD (actuator A) is not required to generate a damping force greater than the normal maximum damping force.

Both damper D and semi-active damper SD (actuator A) generate a damping force that prevents expansion and contraction by receiving the pressure in the rod side chambers 4, 34 and the pressure in the piston side chambers 5, 35 at both axial end faces of the piston 3, 33. Here, to increase the damping force generated by damper D, the pressure-receiving area over which the piston 3 receives the pressure in the rod side chamber 4 and the pressure in the piston side chamber 5 can be increased, or the pressure in the cylinder 1 can be increased, or both can be done.

The piston 3 slides against the inner circumference of the cylinder 1, dividing the inside of the cylinder 1 into a rod side chamber 4 and a piston side chamber 5. Therefore, if the inner diameter of the cylinder 1 is increased, the outer diameter of the piston 3 increases, and the pressure-receiving area of the piston 3 that receives the pressure in the rod side chamber 4 and the pressure in the piston side chamber 5 increases. This makes it possible to increase the damping force generated by the damper D without increasing the pressure in the cylinder 1.

As described above, since the damper D is required to generate a greater damping force than the semi-active damper SD (actuator A), if the inner diameter of the cylinder 1 in the damper D is set to a larger diameter than the inner diameter of the cylinder 31 of the semi-active damper SD (actuator A), the damper D can generate a greater damping force than the semi-active damper SD (actuator A) while suppressing the pressure in the cylinder 1 from increasing. Also, if the inner diameter of the cylinder 1 in the damper D is set to a larger diameter than the inner diameter of the cylinder 31 of the semi-active damper SD (actuator A), the damper D can generate a greater damping force than the semi-active damper SD (actuator A) while suppressing the pressure in the cylinder 1 in the damper D from increasing. Therefore, the requirements for the load resistance and sealing performance of each member constituting the damper D are not as strict, and the product cost of the entire system can be reduced compared to conventional products that must adopt a pressure-resistant structure. Furthermore, it is only necessary to increase the inner diameter of the cylinder 1 of the damper D, and there is no need to increase the inner diameter of the cylinder 31 of the semi-active damper SD.

In addition, instead of increasing the pressure-receiving area over which the piston 3 receives the pressure in the rod side chamber 4 and the pressure in the piston side chamber 5, if the pressure in the cylinder 1 of the damper D is increased to increase the damping force that the damper D can generate, the strength of the damper D can be increased.

When the pressure inside the cylinder 1 is to be increased, for example, the thickness of the cylinder 1 of the damper D can be made thicker than the thickness of the cylinder 31 of the semi-active damper SD (actuator A) to increase the strength of the cylinder 1 and enable it to withstand higher pressure. In this way, by making the thickness of the cylinder 1 of the damper D thicker than the thickness of the cylinder 31 of the semi-active damper SD (actuator A), the strength of the cylinder 1 in the damper D can be improved without increasing the outer diameter compared to the semi-active damper SD (actuator A), and the pressure inside the cylinder 1 can be increased without compromising the mountability of the railway vehicle vibration control device S on a railway vehicle.

In addition, when the pressure inside the cylinder 1 is to be increased, for example, the strength of the material of the cylinder 1, rod 2, and piston 33 of the damper D may be made higher than the strength of the material of the cylinder 31, rod 32, and piston 33 of the semi-active damper SD (actuator A) so that the damper D can withstand high pressure. In this way, if the strength of the material of the cylinder 1, rod 2, and piston 33 of the damper D is made higher than the strength of the material of the cylinder 31, rod 32, and piston 33 of the semi-active damper SD (actuator A), the strength of the damper D is improved without increasing the outer diameter compared to the semi-active damper SD (actuator A), and the pressure inside the cylinder 1 can be increased, and the mountability of the railway vehicle vibration control device S on a railway vehicle can be maintained.

As described above, the railway vehicle vibration control device S of this embodiment has a damper D and either a semi-active damper SD or an actuator A, which are interposed in parallel between the car body B and the bogie T of the railway vehicle, and the required damping force is the sum of the damping forces generated by the damper D and either the semi-active damper SD or the actuator A required when expanding and contracting at a predetermined stroke speed assumed when an earthquake occurs, and the damper D generates a damping force greater than the required damping force when expanding and contracting at a predetermined stroke speed or faster. According to the railway vehicle vibration control device S configured in this way, when expanding and contracting at a predetermined stroke speed assumed when an earthquake occurs, the entire railway vehicle vibration control device S generates the required damping force required when an earthquake occurs only with the damper D, so the semi-active damper SD (actuator A) does not have to share the required damping force when expanding and contracting at a predetermined stroke speed assumed when an earthquake occurs. For this reason, in the railroad vehicle vibration control device S of this embodiment, the damper D is required to generate a larger required damping force than the semi-active damper SD (actuator A) to respond to the occurrence of an earthquake, but the semi-active damper SD (actuator A) is not required to generate a damping force greater than the normal maximum damping force, so there is no need to design the semi-active damper SD (actuator A) to withstand high pressure, and there is also no need to provide the semi-active damper SD (actuator A) with a valve to be used when an earthquake occurs. Therefore, according to the railroad vehicle vibration control device S of this embodiment, while it is possible to generate the required damping force required when an earthquake occurs, there is no need to design the semi-active damper SD or actuator A to withstand high pressure, and the product cost of the entire system can be reduced.

The damper D in the railroad vehicle vibration control device S of this embodiment includes a cylinder 1, a rod 2 inserted into the cylinder 1 so as to be movable in the axial direction, a piston 3 attached to the rod 2 and inserted into the cylinder 1 so as to be movable in the axial direction, dividing the cylinder 1 into a rod side chamber 4 and a piston side chamber 5, a tank 6, a first orifice 8 provided in a first damping passage 7 connecting the rod side chamber 4 and the piston side chamber 5, a second orifice 10 provided in a second damping passage 9 connecting the piston side chamber 5 and the tank 6, a compression side check valve 12 provided in a compression side suction passage 11 connecting the rod side chamber 4 and the tank 6, which allows only the flow of liquid from the tank 6 to the rod side chamber 4, and an extension side check valve 14 provided in an extension side suction passage 13 connecting the piston side chamber 5 and the tank 6, which allows only the flow of liquid from the tank 6 to the piston side chamber 5.

In the railway vehicle vibration control device S having the damper D configured in this way, when the damper D is contracting, the first orifice 8 increases the pressure in the piston side chamber 5 while the compression side check valve 12 opens to make the pressure in the rod side chamber 4 equal to the tank pressure. Therefore, when the damper D is contracting, the pressure-receiving area of the piston 3 that is effective for generating a damping force is the cross-sectional area of the piston 3 itself. Compared to a uniflow type damper, a larger pressure-receiving area can be secured, so that the pressure in the piston side chamber 5 required to generate the same magnitude of damping force is smaller, and the pressure resistance of each component that constitutes the damper D can be reduced, reducing product costs.

In addition, the damper D of this embodiment is provided with an extension side relief valve 16 that is provided in the extension side relief passage 15 that communicates between the rod side chamber 4 and the tank 6 and opens when the pressure in the rod side chamber 4 reaches the valve opening pressure to allow the flow of liquid from the rod side chamber 4 to the tank 6, and a compression side relief valve 18 that is provided in the compression side relief passage 17 that communicates between the piston side chamber 5 and the tank 6 and opens when the pressure in the piston side chamber 5 reaches the valve opening pressure to allow the flow of liquid from the piston side chamber 5 to the tank 6. Therefore, even if the stroke speed during the contraction operation of the damper D becomes faster than the predetermined stroke speed, the extension side relief valve 16 or the compression side relief valve 18 opens, and the pressure in the compression side chamber in the cylinder 1 is prevented from becoming a pressure that the seal members and each member cannot tolerate.

In addition, in the railway vehicle vibration control device S of this embodiment, the damper D and one of the semi-active damper SD or actuator A each have a cylinder 1, 31, a rod 2, 32 inserted axially movably into the cylinder 1, 31, and a piston 3, 33 attached to the rod 2, 32 and inserted axially movably into the cylinder 1, 31 to divide the cylinder 1, 31 into a rod side chamber 4, 34 and a piston side chamber 5, 35, and the inner diameter of the cylinder 1 of the damper D is larger than the inner diameter of the cylinder 31 of the semi-active damper SD or one of the actuators A.

With the railcar vibration control device S configured in this way, the damper D is capable of generating a greater damping force than the semi-active damper SD (actuator A) while suppressing the increase in pressure inside the cylinder 1, so the requirements for the load-bearing capacity and sealing performance of each part of the damper D are not as strict, and the product cost of the entire system can be further reduced compared to conventional products that must adopt a pressure-resistant structure. Also, it is only necessary to increase the inner diameter of the cylinder 1 of the damper D alone, and there is no need to increase the inner diameter of the cylinder 31 of the semi-active damper SD (actuator A).

Furthermore, in the railway vehicle vibration control device S of this embodiment, the damper D and one of the semi-active damper SD or actuator A each have a cylinder 1, 31, a rod 2, 32 inserted axially movably into the cylinder 1, 31, and a piston 3, 33 attached to the rod 2, 32 and inserted axially movably into the cylinder 1, 31 to divide the cylinder 1, 31 into a rod side chamber 4, 34 and a piston side chamber 5, 35, and the thickness of the cylinder 1 of the damper D is thicker than the thickness of the cylinder 31 of the semi-active damper SD or one of the actuators A.

The railroad vehicle vibration damping device S configured in this manner improves the strength of the cylinder 1 in the damper D and increases the pressure inside the cylinder 1 without increasing the outer diameter compared to the semi-active damper SD (actuator A), and does not impair the mountability of the railroad vehicle vibration damping device S on a railroad vehicle.

In the railway vehicle vibration control device S of this embodiment, the damper D and either the semi-active damper SD or the actuator A each have a cylinder 1, 31, a rod 2, 32 inserted axially into the cylinder 1, 31, and a piston 3, 33 attached to the rod 2, 32 and inserted axially into the cylinder 1, 31 to divide the cylinder 1, 31 into a rod side chamber 4, 34 and a piston side chamber 5, 35, and the strength of the material of the cylinder 1, rod 2, and piston 3 of the damper D is higher than the strength of the material of the cylinder 31, rod 32, and piston 33 of the semi-active damper SD or the actuator A.

With the railroad vehicle vibration damping device S configured in this manner, the strength of the damper D can be improved and the pressure inside the cylinder 1 can be increased without increasing the outer diameter compared to the semi-active damper SD (actuator A), and the railroad vehicle vibration damping device S can be mounted on a railroad vehicle without compromising its mountability.

The required damping force is the sum of the damping forces generated by the damper D and either the semi-active damper SD or the actuator A when expanding and contracting at a predetermined stroke speed expected during an earthquake, and as long as the damper D can generate a damping force greater than the required damping force when expanding and contracting at a stroke speed greater than the predetermined stroke speed, the structures of the damper D, the semi-active damper SD and the actuator A in the vibration control device S for railway vehicles are not limited to the specific structures described above and can be modified in design.

Although the preferred embodiment of the present invention has been described in detail above, modifications, variations, and changes are possible without departing from the scope of the claims.

1,31...cylinder, 2,32...rod, 3,33...piston, 4,34...rod side chamber, 5,35...piston side chamber, 6...tank, 7...first damping passage, 8...first orifice, 9...second damping passage, 10...second orifice, 11...compression side suction passage, 12...compression side check valve, 13...rebound side suction passage, 14...rebound side check valve, 15...rebound side relief passage, 16...rebound side relief valve, 17...compression side relief passage, 18...compression side relief valve, A...actuator, B...car body, D...damper, S...railroad vehicle vibration control device, SD...semi-active damper, T...bogie

Claims (5)

A vibration damping device for a railway vehicle, comprising a damper and one of a semi-active damper and an actuator, the damper being interposed in parallel between a carbody and a bogie of the railway vehicle,
The required damping force is the sum of the damping forces generated by the damper and one of the semi-active damper and the actuator when the damper expands and contracts at a predetermined stroke speed expected during an earthquake,
The vibration damping device for a railway vehicle, wherein the damper generates a damping force equal to or greater than the required damping force when the damper expands and contracts at a speed equal to or greater than the predetermined stroke speed. The damper is
A cylinder;
A rod that is inserted into the cylinder so as to be movable in the axial direction;
a piston attached to the rod and inserted into the cylinder so as to be axially movable, the piston dividing the interior of the cylinder into a rod side chamber and a piston side chamber;
Tanks and
a first orifice provided in a first damping passage communicating between the rod side chamber and the piston side chamber;
a second orifice provided in a second damping passage communicating between the piston side chamber and the tank;
a compression side check valve provided in a compression side suction passage communicating between the rod side chamber and the tank, the compression side check valve allowing only a flow of liquid from the tank to the rod side chamber;
2. The vibration damping device for a railway vehicle according to claim 1, further comprising: an extension-side check valve provided in an extension-side suction passage communicating between the piston-side chamber and the tank, the extension-side check valve allowing only a flow of liquid from the tank to the piston-side chamber. Each of the damper and the semi-active damper or the actuator has a cylinder, a rod inserted into the cylinder so as to be axially movable, and a piston attached to the rod, inserted into the cylinder so as to be axially movable, and dividing the inside of the cylinder into a rod side chamber and a piston side chamber,
2. The railcar vibration damping device according to claim 1, wherein an inner diameter of the cylinder of the damper is larger than an inner diameter of a cylinder of one of the semi-active damper and the actuator. Each of the damper and the semi-active damper or the actuator has a cylinder, a rod inserted into the cylinder so as to be axially movable, and a piston attached to the rod, inserted into the cylinder so as to be axially movable, and dividing the inside of the cylinder into a rod side chamber and a piston side chamber,
2. The railcar vibration damping device according to claim 1, wherein the thickness of the cylinder of the damper is greater than the thickness of one of the cylinders of the semi-active damper or the actuator. Each of the damper and the semi-active damper or the actuator has a cylinder, a rod inserted into the cylinder so as to be axially movable, and a piston attached to the rod, inserted into the cylinder so as to be axially movable, and dividing the inside of the cylinder into a rod side chamber and a piston side chamber,
2. The vibration damping device for railway vehicles according to claim 1, wherein the strength of the material of the cylinder, rod and piston of the damper is greater than the strength of the material of the cylinder, rod and piston of the semi-active damper or one of the actuators.

Images