JP4356544B2 - shock absorber - Google Patents

Description

  The present invention relates to a shock absorber used for a suspension of a vehicle or the like.

As shown in FIG. 10, the conventional shock absorber has a movable mass 12 that is floated and supported by a spring 13 in the vertical direction, and includes a resonating portion 7 that is formed integrally with the piston rod 6. The damping force generator 10 is adjusted by the control unit 14 that extends from the lower surface of the movable mass 12 to the damping force generator 10 through the hollow portion of the piston rod 6 with respect to the vibration input. It is known to suppress the vertical vibration of the vehicle body by increasing the vibration damping force by controlling (see, for example, Patent Document 1).
JP 2002-5219 A (page 3, FIG. 1)

However, in the above-described conventional shock absorber, the damping force generating portion is provided in the piston portion, and the damping force is controlled by resonating the movable mass with respect to the vibration input of the vehicle body. Only the damping force generated in the piston part can be controlled, and the damping force generated in the base valve part cannot be controlled. The shock absorber generates a damping force mainly at the piston part on the expansion side, and generates a damping force mainly at the base valve part on the pressure side. However, there is an unsolved problem that the damping force cannot be controlled on the compression side.
Accordingly, the present invention has been made paying attention to the unsolved problems of the above-described conventional example, and an object thereof is to provide a shock absorber capable of mechanically controlling both the compression side and the extension side damping force. It is said.

  In order to achieve the above object, a shock absorber according to the present invention includes a piston built in an inner cylinder, and communicates a piston upper chamber and a piston lower chamber defined by the piston by a first bypass, By communicating the lower piston chamber with a reservoir chamber formed between the outer cylinder and the inner cylinder by a bypass, and controlling the flow of the working fluid in the first and second bypasses by a damping force control means. The damping force control means controls a movable mass that can move in the vertical direction, a switching means that switches between open and closed states of the first and second bypasses in conjunction with the movable mass, and the movable mass. And an urging means for floatingly supporting the frame in the vertical direction.

  According to the present invention, the bypass damping force is controlled by switching the open / close state of the first bypass communicating the piston upper chamber and the piston lower chamber and the second bypass communicating the piston lower chamber and the reservoir chamber. Therefore, it is possible to perform appropriate damping force control by switching the predetermined bypass from the closed state to the open state according to the pressure side and the extension side, and to resonate the movable mass with respect to the vertical vibration on the wheel side. Since the opening / closing state of the bypass is switched by the switching unit interlocked with the movable mass, the damper damping force can be mechanically controlled, and can be realized at a lower cost than the electronically controlled suspension.

Hereinafter, embodiments of the present invention will be described with reference to the drawings.
FIG. 1 is a schematic configuration diagram of a first embodiment of the present invention, in which 1 is a double cylinder type shock absorber, and this shock absorber 1 has a double wall having an inner cylinder 2 and an outer cylinder 3. A cylinder 4 having a structure, a piston rod 5 extending substantially vertically in the inner cylinder 2, a piston 6 connected to the lower end of the piston rod 5 and movable in a substantially axial direction in the inner cylinder 2, and a bottom of the cylinder 4 A base valve 7 is provided, and the cylinder 4 is filled with hydraulic fluid such as oil.
The piston 6 forms a piston upper chamber 8 and a piston lower chamber 9 in the inner cylinder 2, and a reservoir chamber 10 is formed outside the inner cylinder 2.

In this embodiment, the movement of the piston rod 5 in the direction protruding from the cylinder 4 is referred to as “moving to the extension side”, and the movement of the piston rod 5 in the direction of being pushed into the cylinder 4 is “to the pressure side”. "Move".
The upper end portion of the piston rod 5 is connected to the vehicle body 12 via the upper insulator 11, while the eyeball portion 13 at the lower end of the cylinder 4 is attached to a wheel side (not shown).

The shock absorber 1 is provided in the vicinity of the bypass pipe 14 and 15 as a bypass pipe 14 and 15 as a first bypass integrated with the outer cylinder 3 and a predetermined bypass pipe. A damping force control unit 20 that controls the damping force of the shock absorber 1 in response to the unsprung acceleration and frequency is provided.
The bypass pipe 14 communicates with the piston upper chamber 8 and the piston lower chamber 9, and the bypass pipe 15 communicates with the piston lower chamber 9 and the reservoir chamber 10.
The damping force control unit 20 constitutes damping force control means.

FIG. 2 is a configuration diagram illustrating details of the damping force control unit 20. The damping force control unit 20 is installed on the inside of the case 21 filled with hydraulic fluid, the movable mass 22 disposed in the case 21, and above and below the movable mass 22, and elastically floats the movable mass 22. And a spring 23 as a biasing means for supporting.
Further, a passage (orifice) 24 is provided in the movable mass 22 as switching means for controlling a damping force generated by the hydraulic fluid that is driven integrally with the movable mass 22 and passes through the bypass pipes 14 and 15. An attenuator (attenuating orifice) 25 for controlling the displacement of the movable mass 22 is provided. When the movable mass 22 moves upward, the orifice 24 opens a bypass pipe 14 that communicates the piston upper chamber 8 and the piston lower chamber 9, and bypasses the piston lower chamber 9 and the reservoir chamber 10. The pipe 15 is configured to be closed.

Similarly, when the movable mass 22 moves downward, the orifice 24 opens the bypass pipe 15 that communicates the piston lower chamber 9 and the reservoir chamber 10 and communicates the piston upper chamber 8 and the piston lower chamber 9. The bypass pipe 14 is configured to be closed.
When the movable mass 22 is in a substantially neutral position, the bypass pipes 14 and 15 are both closed.

In a general damper, a damping force is generated by a resistance force caused by a pressure difference between both sides of a hole (orifice) in the damper. In the expansion process, a damping force is generated mainly by the resistance generated in the piston valve part, and in the contraction process, a damping force is generated mainly by the resistance generated in the base valve part.
That is, in the expansion process, the damping force can be controlled by controlling the pressure difference between the piston upper chamber 8 and the piston lower chamber 9, and in the contraction process, the piston lower chamber 9 and the reservoir chamber 10 can be controlled. It becomes possible to control the damping force by controlling the pressure difference.

Therefore, in the extension process, by opening the bypass pipe 14, the pressure difference between the piston upper chamber 8 and the piston lower chamber 9 can be reduced and the damping force can be reduced. Further, in the contraction process, by opening the bypass pipe 15, the pressure difference between the piston lower chamber 9 and the reservoir chamber 10 can be reduced and the damping force can be reduced.
At this time, the mass of the movable mass 22 and the spring constant of the spring 23 are set so that the resonance frequency of the resonance system composed of the movable mass 22 and the spring 23 is the same as the unsprung resonance frequency.

  FIG. 3 is a diagram illustrating the relationship between unsprung vibration and movable mass vibration. By setting the mass of the movable mass 22 and the spring constant of the spring 23 so that the resonance frequency of the resonance system is the same as the unsprung resonance frequency, the movable mass displacement is less than the unsprung displacement at the unsprung resonance frequency. Vibrates with a phase delay of 90 °. Since the unsprung speed is 90 degrees out of phase with the unsprung displacement, the movable mass displacement is maximized when the unsprung speed is maximized.

If this characteristic is utilized and the orifice of the spool is opened at the time when the unsprung speed becomes maximum (the piston speed of the shock absorber is maximum), the damping force can be greatly reduced.
In addition, the threshold value of the input that provides a low damping force by opening the spool orifice can be adjusted by a damping coefficient determined by the area and shape of the damping orifice 25 provided in the movable mass 22.

  When driving on road surfaces with large inputs such as single protrusions, irregular roads, and rough roads, the unsprung resonance phenomenon causes unsprung vibration with greater acceleration, speed, and displacement than when driving on other road surfaces. The increase in damping force accompanying the increase is the main factor that deteriorates the ride comfort. Therefore, the damping force is reduced by sensing only a large input unsprung resonance phenomenon different from other road surfaces and steering stability.

That is, at the time of a small input, the displacement of the movable mass 22 is suppressed by the damping force determined by the damping orifice 25 to suppress a low damping force without opening the orifice 24. At the time of a large input, the movable mass 22 is suppressed. The orifice 24 is opened by a large displacement of and is adjusted so as to have a low damping force.
Therefore, it is assumed that the vehicle is currently traveling on a flat road where the vehicle is maintained. In this case, the vertical input from the wheel side is substantially zero, and the base valve 7 is substantially stationary. Therefore, as shown in FIG. 4 (a) when moved to the pressure side and as shown in FIG. 4 (b) when moved to the extension side, the movable mass 22 is kept near the neutral position in the case 21 on both the pressure side and the extension side. Therefore, the bypass pipes 14 and 15 are not opened, and the normal damping force is maintained without becoming a low damping force.

  When the vehicle travels on an irregular road or the like from this state, a large vertical input is applied from the wheel side. In this case, the base valve 7 vibrates substantially vertically, and the movable mass 22 resonates substantially vertically. At this time, when it moves to the pressure side, as shown in FIG. 4C, the movable mass 22 moves downward in the case 21, and the orifice 24 bypasses the piston lower chamber 9 and the reservoir chamber 10. Since the pipe 15 is opened, the damping force is reduced by controlling the pressure difference between the piston lower chamber 9 and the reservoir chamber 10.

On the other hand, when moved to the extension side, the movable mass 22 moves upward in the case 21 and the orifice 24 communicates the piston upper chamber 8 and the piston lower chamber 9 as shown in FIG. Since the bypass pipe 14 is opened, the damping force is reduced by controlling the pressure difference between the piston upper chamber 8 and the piston lower chamber 9.
At this time, the damping force characteristic is as shown in FIG. FIG. 5A shows a damping force characteristic at the time of large input unsprung resonance. A curve indicated by a solid line indicates a damping force characteristic of the shock absorber according to the present invention, and a curve indicated by a broken line indicates a damping force at the time of unsprung resonance. It is a damping force characteristic of a shock absorber in an existing structure without force control. Thus, in the present invention, when the unsprung resonance frequency band is large, the damping force is reduced when the unsprung displacement is minimum, that is, when the unsprung speed is maximized.

  FIG. 5 (b) shows the difference in damping force characteristics due to the difference in input at the time of unsprung resonance. The curve indicated by the solid line is the damping force characteristic at the time of large input, and the characteristic indicated by the broken line is the small input. It is a damping force characteristic at the time. In the present invention, during the unsprung resonance with a small input, the low damping force is not achieved, and the unsprung convergence is promoted. Accordingly, the damping force characteristic at the time of the small input unsprung resonance is substantially the same between the present invention and the existing structure.

As is apparent from FIGS. 5A and 5B, the difference in damping force characteristics between the present invention and the existing structure is as shown in FIG. 5C. In FIG. 5C, the characteristic indicated by the solid line is the unsprung resonance frequency band characteristic of the present invention, and the characteristic indicated by the broken line is the unsprung resonance frequency band characteristic of the existing structure.
As described above, in the shock absorber according to the present invention, the damping force can be reduced when the piston speed is high, and when the piston speed decreases, the damping force returns to the same damping force as the existing structure. It is possible to keep the deterioration of the performance to a minimum, and to prevent the deterioration of the steering stability performance and the riding comfort performance on other road surfaces.

  FIG. 6A is a diagram showing a difference in damping force characteristics due to a difference in input frequency at the same piston speed in the present invention. A solid line indicates a damping force characteristic when an unsprung resonance frequency is input, and a broken line indicates a low value. It is a damping force characteristic at the time of frequency input. As is clear from FIG. 6A, the damping force is not reduced when a low frequency is input, such as when traveling on a wavy road, but is set only when the unsprung resonance frequency band is input. Therefore, it is possible to maintain the steering stability performance and prevent the undulating road from being fully extended.

  FIG. 6B is a diagram showing a difference in damping force characteristics due to a difference in input frequency of the shock absorber in the present embodiment. In FIG. 6B, the curve indicated by the solid line indicates the characteristic in the unsprung resonance frequency band, and the curve indicated by the broken line indicates the characteristic in the low frequency band. As is clear from FIG. 6B, a damping force equivalent to that of normal is generated when a low frequency is input, and the damping force can be reduced when a high frequency and a high piston speed are input. It is possible to perform damping force control that is sensitive to.

  As described above, in the first embodiment, the first bypass that communicates the piston upper chamber and the piston lower chamber and the second bypass that communicates the piston lower chamber and the reservoir chamber are provided. Since the flow of hydraulic fluid in the flow path is controlled, the bypass damping force can be controlled by controlling the pressure difference between the piston upper chamber and the piston lower chamber or the pressure difference between the piston lower chamber and the reservoir chamber, Both damping force control on the extension side can be performed, and the damping force of the damper can be mechanically controlled, so that it can be realized at a lower cost than an electronically controlled suspension.

  Further, a damping force control unit having a movable mass for controlling the bypass damping force, a spring for floatingly supporting the movable mass, an orifice for switching the opening / closing state of the bypass, and an oil passage capable of controlling the displacement of the movable mass is provided on the rod side of the damper. Because it is provided on the outer cylinder side (unsprung side) instead of the (vehicle body side), layout requirements such as shortening the damper stroke and parts characteristics around the upper mount (for example, hood height, upper mount characteristics, bumper rubber characteristics, etc.) The damping force can be controlled without affecting the above.

Further, the mass of the movable mass and the spring constant of the spring that supports the movable mass are set so that the resonance frequency of the resonance system composed of the movable mass and the spring that supports the movable mass substantially matches the unsprung resonance frequency. Therefore, when the unsprung speed becomes maximum, the movable mass displacement can be maximized to open the spool orifice, and the damping force can be greatly reduced.
In addition, the damping coefficient determined by the area and shape of the oil passage provided in the movable mass is adjusted to determine the unsprung input threshold value at which the spool orifice opens, so the damping force is reduced only during large input unsprung resonance. In addition, it is possible to suppress a low damping force at the time of a small input, and it is possible to suppress a deterioration in riding comfort due to an increase in the damping force accompanying a large unsprung vibration.
Furthermore, the damping force is reduced only at the time of a large input, and as the unsprung portion converges, the damping force returns to its original size (becomes higher), so that the deterioration of the unsprung convergence immediately after the large input can be minimized. it can.

Next, a second embodiment of the present invention will be described.
In the second embodiment, the damping force control unit 20 is arranged on the bypass pipe in the first embodiment described above, whereas only the orifice part is arranged on the bypass pipe. is there.
That is, as shown in the detailed configuration diagram of the damping force control unit in FIG. 7, the movable mass 22 provided with the damping orifice 25 outside the bypass pipes 14 and 15 is floatingly supported by the spring 23, and accompanying the unsprung vibration. When the movable mass resonates, the orifice 24 arranged on the bypass pipe is driven.

  Also in the second embodiment, as in the first embodiment described above, when the movable mass 22 moves upward, the bypass piping that connects the piston upper chamber 8 and the piston lower chamber 9 by the orifice 24. 14 is opened, and a bypass pipe 15 communicating the piston lower chamber 9 and the reservoir chamber 10 is closed.

Further, when the movable mass 22 moves downward, the bypass pipe 15 that connects the piston lower chamber 9 and the reservoir chamber 10 is opened by the orifice 24 and the bypass that connects the piston upper chamber 8 and the piston lower chamber 9 is opened. The pipe 14 is configured to be closed.
Thereby, in the expansion process, by opening the bypass pipe 14, the pressure difference between the piston upper chamber 8 and the piston lower chamber 9 can be reduced, and the damping force can be reduced. In the contraction process, the bypass pipe 15 can be reduced. By opening, the pressure difference between the piston lower chamber 9 and the reservoir chamber 10 can be reduced to reduce the damping force.
Again, when the unsprung speed is maximized by setting the mass of the movable mass and the spring constant of the spring that supports the movable mass so that the resonance frequency of the movable mass and the unsprung resonance frequency are substantially the same. The spool orifice can be opened with the maximum movable mass displacement, and the damping force can be greatly reduced.

In addition, the damping coefficient determined by the area and shape of the oil passage provided in the movable mass is adjusted to determine the unsprung input threshold value at which the spool orifice opens. It is possible to reduce and suppress a low damping force at the time of a small input, and it is possible to suppress a deterioration in riding comfort due to an increase in the damping force accompanying a large unsprung vibration.
Thus, in the said 2nd Embodiment, the effect similar to 1st Embodiment mentioned above can be acquired.

Next, a third embodiment of the present invention will be described.
In the third embodiment, the opening and closing timing of the bypass pipe is shifted with respect to the unsprung speed maximum time.
That is, in the first and second embodiments described above, the mass of the movable mass 22 and the spring 23 are set so that the resonance frequency of the resonance system composed of the movable mass 22 and the spring 23 is the same as the unsprung resonance frequency. In the third embodiment, the bypass pipes 14 and 15 are opened and closed by setting the spring constant so that the movable mass displacement is maximized when the unsprung speed is maximized. Then, the time when the unsprung speed becomes maximum and the time when the movable mass displacement becomes maximum are set to be shifted.

Specifically, the mass of the movable mass 22 and the spring constant of the spring 23 are set so that the resonance frequency of the movable mass 22 is lower than the unsprung resonance frequency.
As a result, as shown in FIG. 8, the phase of the movable mass 22 at the unsprung resonance frequency becomes 90 ° or more, and the opening and closing timing of the bypass pipes 14 and 15 is delayed with respect to the maximum unsprung speed, that is, the damping force Reduction timing can be delayed.

By utilizing this phenomenon, characteristics as shown in FIG. 9 can be obtained. FIG. 9A is a diagram showing the difference in damping force characteristics due to the difference in the resonance frequency of the movable mass, and the curve shown by the solid line is the case where the resonance frequency of the movable mass is set lower than the unsprung resonance frequency. The damping force characteristic, the curve indicated by the broken line, is the damping force characteristic when the unsprung resonance frequency and the resonance frequency of the movable mass are set equal.
As described above, by delaying the opening / closing timing of the bypass pipes 14 and 15 with respect to the maximum unsprung speed time, the damping force at the time of maximum spring displacement is reduced as compared with the case where the opening / closing timing matches the maximum unsprung time. can do.

  Incidentally, the vehicle body input is the sum of the shock absorber input and the suspension spring (suspension spring) input. FIG. 9B is a diagram illustrating the vehicle body input characteristics when the unsprung resonance frequency and the resonance frequency of the movable mass are set to be equal, and the opening / closing timing of the bypass pipe is matched with the maximum unsprung speed time. In FIG. 9B, the broken line indicates the suspension spring input, the alternate long and short dash line indicates the shock absorber input, and the solid line indicates the total vehicle body input (shock absorber input + suspension spring input).

The suspension spring input becomes maximum at the time when the unsprung displacement becomes maximum. Part A shows the maximum peak of the total vehicle body input. In order to reduce the total vehicle body input, it is only necessary to reduce the shock absorber input at the time when the unsprung displacement is maximum.
FIG. 9C is a diagram showing the vehicle body input characteristics when the resonance frequency of the movable mass is set lower than the unsprung resonance frequency and the opening / closing timing of the bypass pipe is delayed with respect to the maximum unsprung speed time. . In FIG. 9C, the broken line indicates the suspension spring input, the alternate long and short dash line indicates the shock absorber input, and the solid line indicates the total vehicle body input (shock absorber input + suspension spring input). Further, part B shows the maximum peak of the total vehicle body input.

  As is clear from comparison between the A part and the B part in FIGS. 9B and 9C, when the opening / closing timing of the bypass pipe is delayed with respect to the maximum unsprung speed, the opening / closing timing of the bypass pipe Since the damping force at the maximum unsprung displacement can be reduced compared to the case where is matched with the maximum time of unsprung speed, the maximum value of the vehicle body input, which is the sum of the shock absorber input and the suspension spring input Can be reduced.

  As described above, in the third embodiment, by setting the resonance frequency of the movable mass to be lower than the unsprung resonance frequency, the resonance phase of the movable mass is delayed with respect to the unsprung resonance. The maximum value of the vehicle body input, which is the sum of the input and the suspension spring input, can be reduced and attenuated, and both the reduction in ride comfort vibration and convergence can be achieved.

  In the third embodiment, the case where the resonance frequency of the movable mass is set lower than the unsprung resonance frequency has been described. However, the present invention is not limited to this, and the movable mass is provided in the movable mass. By increasing the damping force of the oil passage (damping orifice), the phase of the movable mass at the unsprung resonance frequency is set to 90 ° or more, and the opening / closing timing of the bypass pipe is delayed with respect to the maximum time of the unsprung speed. Also good.

It is a schematic structure figure showing an embodiment of the present invention. It is a block diagram which shows the detail of the damping force control part in 1st Embodiment. It is a figure which shows the relationship between the unsprung vibration and movable mass vibration in 1st Embodiment. It is a figure explaining the operation | movement in 1st Embodiment. It is a figure which shows the damping force characteristic by the difference in the magnitude | size of an input. It is a figure which shows the damping force characteristic by the difference in input frequency. It is a block diagram which shows the detail of the damping force control part in 2nd Embodiment. It is a figure which shows the relationship between the unsprung vibration and movable mass vibration in 2nd Embodiment. It is a figure explaining the effect by change of damping force reduction timing. It is a schematic block diagram of the shock absorber in a prior art.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Shock absorber 2 Inner cylinder 3 Outer cylinder 4 Cylinder 5 Piston rod 6 Piston 7 Base valve 8 Piston upper chamber 9 Piston lower chamber 10 Reservoir chamber 11 Upper insulator 12 Car body 13 Eyepiece 14 Bypass piping 15 Bypass piping 20 Damping force control section 21 Case 22 Movable mass 23 Spring 24 Orifice 25 Damping orifice

Claims (6)

In a multi-cylinder shock absorber that is installed between the vehicle body side and the wheel side and is composed of an outer cylinder and an inner cylinder that is installed in the outer cylinder,
The piston is formed between the piston built in the inner cylinder, the first bypass communicating the piston upper chamber and the piston lower chamber partitioned by the piston, the piston lower chamber, the outer cylinder, and the inner cylinder. A second bypass communicating with the reservoir chamber, and a damping force control means for controlling the damping force by controlling the flow of the working fluid in the first and second bypasses, the damping force control means comprising: A movable mass movable in the vertical direction, switching means for switching the open and closed states of the first and second bypasses in conjunction with the movable mass, and an urging means for floatingly supporting the movable mass in the vertical direction A shock absorber characterized by comprising.   The mass of the movable mass and the spring constant of the biasing means are set so that the resonance frequency of the resonance system composed of the movable mass and the biasing means is equal to the unsprung resonance frequency, and the movable mass is 2. The first and second bypasses are configured such that the first or second bypass is opened at a position of an antinode when the resonance occurs. Shock absorber described in 1.   The shock absorber according to claim 1, wherein the shock absorber is configured to delay the phase of resonance of the movable mass with respect to unsprung resonance.   The switching means switches the first bypass from the closed state to the open state when the piston moves to the extension side, and changes the second bypass from the closed state to the open state when the piston moves to the pressure side. The shock absorber according to any one of claims 1 to 3, wherein the shock absorber is configured to be switched.   The damping force control means reduces the damping force by controlling a pressure difference between two chambers communicated with a bypass that is switched from a closed state to an open state by the switching means at a predetermined frequency. The shock absorber according to any one of claims 1 to 4, wherein the shock absorber is configured as described above.   At a predetermined frequency, either the bypass communicating the piston upper chamber and the piston lower chamber or the bypass communicating the piston lower chamber and the reservoir chamber is switched from the closed state to the open state. A shock absorber configured to reduce a damping force by controlling a pressure difference between two chambers communicated by a bypass switched to an open state.

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