JP2009243636A - Damping force adjustable shock absorber and suspension control device using same - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To prevent excessive increase of damping force by inhibiting a rise of pressure in a back pressure chamber in response to high frequency vibration in a pilot (back pressure) type damping force adjustable hydraulic shock absorber. <P>SOLUTION: The damping force adjustable shock absorber generates damping force by controlling flow of oil between an annular oil passage 21 and a reservoir 4 generated by a slide movement of a piston in a cylinder by a pilot (back pressure) type damping valve 27 and a pressure control valve 28 (solenoid control valve). The damping force is directly generated by the pressure control valve 28 and valve-opening pressure of the damping valve 27 is controlled by adjusting inner pressure of the back pressure chamber 53. Since resistant force against movement in a valve-opening direction is larger than resistant force against movement in a valve-closing direction in a valve body 58 of the pressure control valve 28, valve close delay to unsprung resonance (high frequency vibration) of a suspension device occurs, a valve opening state is maintained to continuous input, and damping force drops. Consequently, deterioration of ride comfort due to excessive rise of damping force can be prevented. <P>COPYRIGHT: (C)2010,JPO&INPIT

Description

  The present invention relates to a damping force adjusting type shock absorber attached to a suspension device of a vehicle such as an automobile, and a suspension control device using the same.

  In general, a damping force adjusting type shock absorber mounted on an automobile suspension device is slidably fitted with a piston connected to a piston rod in a cylinder in which oil is sealed, and the inside of the cylinder is divided into two chambers. The flow of oil generated by the sliding of the piston in the cylinder is controlled by a damping force generating mechanism including an orifice, a disk valve, etc. to generate a damping force, and a flow control valve, a pressure control valve, etc. are used. The damping force is adjusted by changing the flow resistance of the damping force generation mechanism.

In this type of damping force adjusting type shock absorber, as described in Patent Document 1, for example, a back pressure chamber is formed at the back of a disc valve which is a damping force generating mechanism, and this back pressure chamber is fixed to a fixed orifice. There are some which are connected to the upstream cylinder chamber via a pressure control valve (solenoid valve) and connected to the downstream cylinder chamber via a pressure control valve (solenoid valve).
JP 2001-12534 A

  With this configuration, the flow resistance of oil can be adjusted directly by the pressure control valve, and the opening pressure of the disc valve can be adjusted by adjusting the internal pressure of the back pressure chamber, so that the adjustment range of the damping force characteristic can be widened. can do.

  However, the damping force adjustment type shock absorber described in Patent Document 1 has the following problems. For example, when mounted on a suspension control device of a vehicle such as an automobile and executing damping force control by a control signal from a controller according to the running state of the vehicle, with respect to unsprung resonance (high frequency vibration) of the suspension device, Response (opening) delay may occur due to the inertia of the plunger of the pressure control valve (solenoid valve), the pressure in the back pressure chamber will rise, the damping force will increase excessively, and the riding comfort may deteriorate.

  The present invention has been made in view of the above points, and a damping force adjustment type shock absorber capable of preventing an excessive increase in damping force by suppressing an increase in pressure in the back pressure chamber against high frequency vibration. And it aims at providing the suspension control apparatus using the same.

In order to solve the above-mentioned problems, the present invention provides a cylinder in which a fluid is sealed, a piston slidably fitted in the cylinder, one end connected to the piston, and the other end outside the cylinder. A damping force adjustment comprising: a piston rod extended to the inside of the cylinder; and a pressure control valve capable of adjusting a valve opening pressure by controlling a flow of fluid generated by sliding of the piston in the cylinder to generate a damping force. In the type shock absorber,
The valve body of the pressure control valve has greater resistance to movement in the valve closing direction than resistance to movement in the valve opening direction, and a valve closing delay occurs with respect to a predetermined high frequency input. And

  According to the present invention, for a predetermined high frequency input, for example, vibration of the unsprung resonance frequency of the suspension device, the valve closing delay occurs in the pressure control valve, the valve open state is maintained for the continuous input, and the damping force Therefore, it is possible to prevent the ride comfort from being deteriorated due to an excessive increase in damping force.

Hereinafter, an embodiment of the present invention will be described in detail with reference to the drawings.
As shown in FIG. 2, the damping force adjusting hydraulic shock absorber 1 (fluid pressure shock absorber) according to the present embodiment has a double cylinder structure in which an outer cylinder 3 is provided outside the cylinder 2. A reservoir 4 is formed between the outer cylinder 3 and the outer cylinder 3. A piston 5 is slidably fitted in the cylinder 2, and the inside of the cylinder 2 is defined by the piston 5 as two chambers, a cylinder upper chamber 2A and a cylinder lower chamber 2B. One end of a piston rod 6 is connected to the piston 5 by a nut 7, and the other end side of the piston rod 6 passes through the cylinder upper chamber 2 </ b> A and is a rod guide attached to the upper ends of the cylinder 2 and the outer cylinder 3. 8 and the oil seal 9 are extended to the outside of the cylinder 2. A base valve 10 that partitions the cylinder lower chamber 2 </ b> B and the reservoir 4 is provided at the lower end of the cylinder 2.

  The piston 5 is provided with oil passages 11 and 12 for communicating between the cylinder upper and lower chambers 2A and 2B. The oil passage 11 is provided with a check valve 13 that allows only fluid to flow from the cylinder lower chamber 2B side to the cylinder upper chamber 2A side, and the oil passage 12 has a cylinder upper chamber 2A side. A disk valve 14 is provided that opens when the pressure of the oil liquid reaches a predetermined pressure and relieves it to the cylinder lower chamber 2B side.

  The base valve 10 is provided with oil passages 15 and 16 that allow the cylinder lower chamber 2 </ b> B and the reservoir 4 to communicate with each other. The oil passage 15 is provided with a check valve 17 that allows only fluid to flow from the reservoir 4 side to the cylinder lower chamber 2B side, and the oil passage 16 has oil in the cylinder lower chamber 2B side. A disk valve 18 is provided that opens when the liquid pressure reaches a predetermined pressure and relieves it to the reservoir 4 side. An oil liquid is sealed in the cylinder 2, and an oil liquid and a gas are sealed in the reservoir 4.

  An outer tube 20 is fitted on the cylinder 2 at both upper and lower ends via seal members 19, and an annular oil passage 21 is formed between the cylinder 2 and the outer tube 20. The annular oil passage 21 is communicated with the cylinder upper chamber 2 </ b> A by an oil passage 22 provided on a side wall near the upper end portion of the cylinder 2. A small-diameter opening 23 is provided on the side wall of the outer tube 20, and a large-diameter opening 24 is provided substantially concentrically with the opening 23 on the side wall of the outer cylinder 3. A damping force generating mechanism 25 is attached to 24.

  The damping force generation mechanism 25 will be described with reference to FIG. As shown in FIG. 1, one end of a cylindrical case 26 is inserted into the opening 24 and fixed by welding. A valve unit 30 in which a pilot type (back pressure type) damping valve 27 and a pressure control valve 28 (solenoid control valve) are integrated is inserted into the case 26 and fixed by a nut 31.

  The valve unit 30 includes a solenoid case 32 that is fixed to the case 26 by a nut 31. A guide bore 33 is formed along the axial direction at the outer end of the solenoid case 32, and a plunger 34 is slidably guided in the guide bore 33. Further, a coil 35, a plunger spring 36 and a core 37 are accommodated. These are fixed by attaching the base 38 to the solenoid case 32 with a nut 39. An adjustment screw 40 for adjusting the spring force of the plunger spring 36 is attached to the base 38. The coil 35 is connected to a lead wire 35A for energization and extends to the outside.

  A passage bore 41 concentric with the guide bore 33 is formed at the inner end of the solenoid case 32, and the guide bore 33 and the passage bore 41 communicate with each other via a small diameter port 42. A bottomed cylindrical guide member 43 and a stepped cylindrical valve member 44 are disposed in this order at the inner end of the solenoid case 32, and a small diameter portion of the valve member 44 is inserted through the bottom of the guide member 43. The threaded portion at the tip is screwed into the passage bore 41 so that they are integrally coupled. The large diameter portion of the stepped cylindrical passage member 45 is fitted into the large diameter portion of the valve member 44, and the small diameter portion of the passage member 45 is inserted into the port member 46 welded to the opening 23 of the outer tube 20. The chamber 47 inside the valve member 44 is communicated with the annular oil passage 21 via the passage member 45. A chamber 48 formed around the valve unit 30 in the case 26 communicates with the reservoir 4.

  A plurality of oil passages 49 communicating with the chamber 47 are provided at the bottom portion of the valve member 44, and an annular valve seat 50 projects from the outer end surface of the bottom portion on the outer peripheral side of the oil passage 49. Between the valve member 44 and the guide member 43, the inner peripheral portion of a plurality of stacked disc valves 51 (main valve) is clamped, and the outer peripheral portion of the disc valve 51 is seated on the valve seat 50. An annular seal member 52 is fixed to the back surface of the disc valve 51, and the seal member 52 is fitted in a liquid-tight and slidable manner on the inner peripheral surface of the cylindrical portion of the guide member 43, so that the guide A back pressure chamber 53 is formed inside the member 43. Then, the disk valve 51 receives the pressure of the oil liquid in the oil passage 49 and bends and separates (opens) from the valve seat 50 to directly connect the oil passage 49 to the chamber 48. At this time, the disc valve 51 and the back pressure chamber 53 form a pilot type (back pressure type) damping valve, and the internal pressure of the back pressure chamber 53 acts in the valve closing direction of the disc valve 51. . The axial oil passage 55 in the small diameter portion of the valve member 44 has one end communicating with the chamber 47 through the fixed orifice 56, the other end communicating with the passage bore 41, and the back through the radial oil passage 57. The pressure chamber 53 communicates.

  A valve body 58 that opens and closes the port 42 is attached to the distal end portion of the plunger 34 so as to be movable in the axial direction. The valve body 58 is a valve spring 59 interposed between the valve body 58 and the plunger 34. (Coil spring) is pressed against the seat surface 42A (valve seat) around the port 42 to close the port 42. The port 42 and the valve body 58 form the pressure control valve 28. The valve body 58 opens away from the seat surface 42A when the oil pressure in the port 42 reaches a predetermined pressure. The valve opening pressure is adjusted in accordance with the spring force of the plunger spring 36 and the thrust of the solenoid, that is, the energization current to the coil 35. A chamber 33 </ b> A formed in the guide bore 33 by the distal end portion of the plunger 34 is communicated with the chamber 48 through an oil passage 60 formed in the solenoid case 32.

  As shown in FIG. 3, the plunger 34 is formed in a stepped cylindrical shape having a small diameter portion 34 </ b> A that abuts the valve body 58 on one end side, penetrates in the axial direction in the center portion, and the chamber 33 </ b> A in the guide bore 33 A guide passage 61 that communicates with a chamber 33B formed in the back portion of the plunger 34 is provided. A tapered chamfer 61 </ b> A is formed at the peripheral edge of the opening on the small diameter portion 34 </ b> A side of the guide passage 61. The valve body 58 has a small diameter portion 58A inserted into the guide passage 61 of the plunger 34 with a predetermined gap, a medium diameter contact portion 58B that contacts the tip of the small diameter portion 34A of the plunger 34, and a large valve spring 59. An annular seat 58D is formed on the end of the spring receiving portion 58C that is attached to and detached from the seat surface 42A around the port 42. A tapered portion 58E facing the chamfered portion 61A of the guide passage 61 is formed at the base portion of the small diameter portion 58A of the valve body 58, and as shown in FIG. A predetermined gap is formed between the chamfered portion 61A and the tapered portion 58E when contacting the small diameter portion 34A. As shown in FIG. 4, a plurality of orifice grooves 34 </ b> B extending in the radial direction (four at equal intervals in the circumferential direction in the illustrated example) are formed at the distal end portion of the small diameter portion 34 </ b> A of the plunger 34. Has been.

  As shown in FIG. 5, when the contact portion 58B of the valve body 58 contacts the small diameter portion 34A of the plunger 34, the chambers 33A and 33B at both ends of the plunger 34 are communicated with each other via the orifice groove 34B. It has become so. In the illustrated example, the orifice grooves 34B are radially arranged at four locations at equal intervals in the circumferential direction.

  The area of the flow path (variable orifice flow path) communicating between the chambers 33A and 33B on both sides of the plunger 34 in the guide bore 33 via the guide passage 61 is the small diameter portion 34A of the plunger 34 and the contact portion 58B of the valve body 58. 11, assuming that the flow path area of the gap between the small diameter portion 58A of the valve body 58 and the guide passage 61 is E and the flow path area of the orifice groove 34B is F. FIG. As shown, the maximum is the flow path area E, and decreases as the gap C decreases, and the minimum is the flow path area F. The relationship between the attenuation coefficient c1 and the gap C due to the flow path area is as shown in FIG. The initial value of the gap C can be adjusted by changing the spring force of the plunger spring 36 with the adjusting screw 40.

  When the valve body 58 of the pressure control valve 28 starts to open, that is, when the seat portion 58D starts to separate from the seat surface 42A, the gap C is large, so that the oil liquid with respect to the movement of the valve body 58 in the valve opening direction The resistance force due to is small, and it is easy to open the valve. On the other hand, when the valve is closed after the valve is opened, the gap C is small, the resistance by the oil to the movement of the valve body 58 in the valve closing direction is large, and the valve is difficult to close.

  Next, a vibration system model of the pressure control valve 28 is shown in FIG. 10, M1 is the mass of the plunger 34, M2 is the mass of the valve body 58, k1 is the elastic coefficient of the plunger spring 36, k2 is the elastic coefficient of the valve spring 59, x1 is the displacement of the plunger, and x2 is the displacement of the valve body 58. , C1 represents a damping coefficient for the displacement of the plunger 34, c2 represents a damping coefficient for the displacement of the valve body 58, and F represents an external force acting on the valve body 58. The attenuation coefficient c1 changes according to the gap C as shown in FIG.

The equations of motion relating to the displacement x1 of the plunger 43 and the displacement x2 of the valve body 58 can be expressed by the following equations (Equation 1 and Equation 2).
M1 · x1 ″ = − k1 · x1−k2 · (x1−x2) −c1 · x1′−c2 · (x1′−x2 ′) (Equation 1)
M2 · x2 ″ = k2 · (x1−x2) + c2 · (x1′−x2 ′) + F (Expression 2)

  The valve body 58 of the pressure control valve 28 has a large displacement (amplitude) without causing a delay in closing the valve with respect to a low-frequency input lower than the unsprung resonance frequency of the suspension device of the vehicle. For high-frequency input near the lower resonance frequency, valve closing delay occurs, and the amount of displacement (amplitude) becomes small. Regardless of the valve opening pressure (current applied to the coil 35), the valve opens for continuous input. The state is maintained.

Next, the operation of the present embodiment configured as described above will be described.
The damping force adjusting hydraulic shock absorber 1 is connected to the suspension device of a vehicle such as an automobile, the cylinder 2 side is connected to the lower spring side, the piston rod 6 side is connected to the upper spring side, and the lead wire of the coil 35 is connected. 35A is connected to a controller (not shown) and attached to the suspension control device.

  During the extension stroke of the piston rod 6, the check valve 12 of the piston 5 is closed by the movement of the piston 5 in the cylinder 2, and the oil liquid on the cylinder upper chamber 2 </ b> A side is pressurized before the disk valve 14 is opened. Then, the fluid flows from the passage member 45 to the chamber 47 of the damping force generation mechanism 25 through the oil passage 22 and the annular oil passage 21. Before the disc valve 51 of the main valve 27 is opened, the oil passes from the chamber 47 through the fixed orifice 56, the axial oil passage 55, the passage bore 41 and the port 42, and the valve body 58 of the pressure control valve 28. Is opened to flow to the chamber 33A in the guide bore 33, and further to the reservoir 4 through the oil passage 60 and the chamber 48. When the pressure in the chamber 47 reaches the valve opening pressure of the disc valve 51, the disc valve 51 is opened, and the oil liquid flows directly from the chamber 47 to the chamber 48.

  At this time, the oil corresponding to the movement of the piston 5 opens the check valve 17 of the base valve 10 from the reservoir 4 and flows into the cylinder lower chamber 2B. When the pressure in the cylinder upper chamber 2A reaches the valve opening pressure of the disk valve 14 of the piston 5, the disk valve 14 is opened, and the pressure in the cylinder upper chamber 2A is relieved to the cylinder lower chamber 2B. Prevent excessive pressure rise of 2A.

  During the contraction stroke of the piston rod 6, the check valve 13 of the piston 5 is opened by the movement of the piston 5 in the cylinder 2, the check valve 17 of the oil passage 15 of the base valve 10 is closed, and the disc valve 18 is opened. Before, the fluid in the piston lower chamber 2B flows into the cylinder upper chamber 2A, and the amount of fluid that has entered the cylinder 2 from the piston rod 6 passes from the cylinder upper chamber 2A through the same path as in the extension stroke. Flows through to the reservoir 4. When the pressure in the cylinder lower chamber 2B reaches the valve opening pressure of the disk valve 18 of the base valve 10, the disk valve 18 is opened, and the pressure in the cylinder lower chamber 2B is relieved to the reservoir 4, thereby Prevent excessive pressure rise of 2B.

  As a result, during the expansion / contraction stroke of the piston rod 6, before the main valve 27 is opened (piston speed low speed region), a damping force is generated by the fixed orifice 56 and the pressure control valve 28, and after the main valve 27 is opened. In the (piston speed high speed region), a damping force is generated according to the opening degree. The damping force can be directly controlled regardless of the piston speed by adjusting the valve opening pressure of the pressure control valve 28 by the energization current to the coil 35. At this time, since the internal pressure of the back pressure chamber 53 is adjusted by the valve opening pressure of the pressure control valve 28, the valve opening pressure of the main valve 27 can be adjusted simultaneously, and the adjustment range of the damping force characteristic can be widened. it can.

  Here, the resistance force of the pressure control valve 28 to the movement of the valve body 58 changes depending on the change in the flow path area of the gap C. When the valve body 58 is opened due to the pressure of the oil liquid from the port 42, first, the valve body 58 is retracted, the clearance C between the plunger 34 is reduced, and the valve body 58 comes into contact with the plunger 34 (see FIG. 5), and thereafter, the valve body 58 and the plunger 34 are moved backward together (see FIG. 6). At this time, when the valve body 58 is opened, the clearance C is large, and the resistance force by the oil to the movement of the valve body 58 in the valve opening direction is small, so that the valve body 58 is easy to open. Since the clearance C is small and the resistance force against the movement of the valve body 58 in the valve closing direction is large, it is difficult to close the valve. Further, due to the characteristics of the vibration system represented by the model shown in FIG. 10 and Formulas 1 and 2, the valve body 58 has a large displacement (amplitude) without causing a valve closing delay with respect to a low frequency input, A damping force corresponding to the valve opening pressure (current applied to the coil 35) is generated. However, for a high-frequency input near the unsprung resonance frequency, a valve closing delay occurs, and the displacement (amplitude) decreases. Since the valve open state is maintained with respect to the continuous input, the damping force is reduced regardless of the valve opening pressure (the current applied to the coil 35).

  For example, as shown in FIG. 13A, when the input frequency to the piston rod 6 is 1 Hz (low frequency) (in the damping force generation mechanism 25, the oil liquid is in the same flow path with respect to the expansion and contraction of the piston rod 6. Therefore, the input frequency to the valve body 58 of the pressure control valve 28 is 2 Hz.) As shown in FIG. 13B, the valve body 58 can follow the continuous input, The flying height Y1 from the seat surface 42A when the valve is closed is sufficiently small.

  On the other hand, as shown in FIG. 14A, when the input frequency to the piston rod 6 is 20 Hz (high frequency) (in the damping force generation mechanism 25, the oil liquid flows in the same flow for the expansion and contraction of the piston rod 6). Since the flow is circulated, the input frequency of the pressure control valve 28 to the valve body 58 is 40 Hz.) As shown in FIG. On the other hand, the flying height Y2 from the seat surface 42 when the valve is closed increases, and the valve open state is maintained.

  Thereby, the relationship between the input frequency and the flying height from the seat surface 42A when the valve body 58 is closed is as shown in FIG. 15, and the flying height when the valve is closed is small for low-frequency input. In addition, a predetermined damping force can be generated according to the energization current to the coil 35, and the floating amount when the valve is closed is large for the high frequency input, and the valve open state is maintained for the continuous input. Therefore, the damping force can be sufficiently reduced regardless of the energization current to the coil 35.

  As a result, the damping force is controlled by controlling the valve opening pressure of the pressure control valve 28 by the control current from the controller in the normal state where the unsprung vibration of the suspension device due to the road surface input is lower than the unsprung resonance frequency. be able to. When the input frequency from the road surface increases and reaches the vicinity of the unsprung resonance frequency of the suspension device, the valve body 58 of the pressure control valve 28 causes a valve closing delay with respect to continuous input and maintains the valve open state. Thus, the damping force becomes sufficiently small, so that vibration under the spring can be absorbed and transmission to the spring (vehicle body side) can be cut off, and the ride comfort can be improved. In this way, the damping force can be reduced with respect to the unsprung resonance regardless of the control current from the controller, and appropriate damping force control can be performed.

  Next, a modification of the pressure control valve 28 of the above embodiment will be described with reference to FIGS. In addition, the same code | symbol is attached | subjected to the same part with respect to the said embodiment, and only a different part is demonstrated in detail.

  7 to 9, a cylindrical portion 62 is formed at the tip of the plunger 34 instead of the small diameter portion 34 </ b> A, and a stepped columnar valve body 58 is slidable in the cylindrical portion 62. A leaf spring-like valve spring 63 is interposed between the bottom portion of the cylindrical portion 62 and the valve body 58. The valve body 58 is provided with a passage 64 penetrating in the axial direction. An orifice groove 65 extending in the radial direction is formed at the bottom of the cylindrical portion 62, and the passage 64 and the guide passage 61 of the plunger 34 are always in communication with each other by the orifice groove 65. In the illustrated example, four orifice grooves 65 are arranged radially at equal intervals. As shown in FIG. 8, the valve spring 63 is formed in a substantially cross shape in which leg portions 63 </ b> B that contact the bottom of the cylindrical portion 62 are extended in four directions at equal intervals from the central portion 63 </ b> A that contacts the valve body 58. As shown in FIG. 9, the valve body 58 is arranged so as not to close the passage 64 in a state where the valve body 58 has moved to the bottom side of the cylindrical portion 62.

  With this configuration, the same operational effects as those of the above embodiment can be obtained. Further, by fitting the valve body 58 in the cylindrical portion 62 and using the valve spring 63 as a leaf spring, it is possible to improve the assembling property and reduce the manufacturing cost.

  In the above-described embodiment and the modification thereof, as an example, the pressure control valve 28 is used to control the pressure in the back pressure chamber 53 according to the control current to the coil 35. However, the present invention can be similarly applied to a device that controls the pressure control valve by other driving means.

  In the above-described embodiment and its modification, the valve unit 30 in which the damping valve 27 and the pressure control valve 28 are integrated is disposed in the case 26 on the side of the cylinder 2, and the annular oil passage 21, the reservoir 4, The damping fluid is generated by controlling the flow of oil between the valve unit 30 and the valve unit 30 is disposed on the piston 5 or the base valve 10 to appropriately control the flow of oil in the oil passage. A damping force may be generated.

  Furthermore, in the above-described embodiment and its modification, the hydraulic shock absorber that generates a damping force by controlling the flow of the oil liquid has been described. However, the present invention is not limited to this, and other fluids such as gas. The present invention can be similarly applied to a device that generates a damping force by controlling the flow of the.

It is a longitudinal cross-sectional view which expands and shows the damping force generation mechanism of the damping force adjustment type hydraulic buffer which concerns on one Embodiment of this invention. 1 is a longitudinal sectional view of a damping force adjusting hydraulic shock absorber according to an embodiment of the present invention. It is a longitudinal cross-sectional view which expands and shows the pressure control valve of the damping force generation mechanism of FIG. It is a longitudinal cross-sectional view by the AA line of FIG. FIG. 4 is a longitudinal sectional view showing a state in which the valve body is retracted and abuts against a plunger in the pressure control valve of FIG. 3. FIG. 4 is a longitudinal sectional view showing a state in which the valve body is retracted to contact the plunger and the plunger is further retracted in the pressure control valve of FIG. 3. It is a longitudinal cross-sectional view which shows the modification of the pressure control valve of FIG. It is a top view which shows the valve spring of the pressure control valve of FIG. FIG. 8 is a longitudinal sectional view showing a state in which the valve body is retracted in the pressure control valve of FIG. 7. It is the schematic which modeled the vibration system of the plunger and valve body of the pressure control valve of FIG. It is a graph which shows the relationship between the clearance gap C of the plunger and valve body of the pressure control valve of FIG. 3, and its flow-path area. It is a graph which shows the relationship between the clearance gap C of the plunger and valve body of the pressure control valve of FIG. 3, and its attenuation coefficient. It is a graph which shows the displacement of the valve body with respect to the low frequency input of the pressure control valve of FIG. It is a graph which shows the displacement of the valve body with respect to the high frequency input of the pressure control valve of FIG. In the pressure control valve of FIG. 3, it is a figure which shows the relationship between an input frequency and the flying height from the seat surface at the time of valve closing of a valve body.

Explanation of symbols

  1 Damping force adjusting hydraulic shock absorber (damping force adjusting shock absorber), 2 cylinder, 5 piston, 6 piston rod, 28 pressure control valve, 58 valve body

Claims (20)

A cylinder filled with fluid, a piston slidably fitted in the cylinder, a piston rod having one end connected to the piston and the other end extending outside the cylinder, In a damping force adjusting type shock absorber provided with a pressure control valve capable of adjusting a valve opening pressure by generating a damping force by controlling a flow of fluid generated by sliding of the piston,
The valve body of the pressure control valve has greater resistance to movement in the valve closing direction than resistance to movement in the valve opening direction, and a valve closing delay occurs with respect to a predetermined high frequency input. Damping force adjustable shock absorber. A main valve that generates a damping force by controlling a flow of fluid generated by sliding of the piston in the cylinder; and a back pressure chamber that applies an internal pressure to the main valve in a valve closing direction. A part of the pressure is introduced into the back pressure chamber, the opening of the main valve is controlled by the internal pressure of the back pressure chamber, and the internal pressure of the back pressure chamber is adjusted by the pressure control valve. The damping force adjusting type shock absorber according to 1. 3. The damping force according to claim 1, wherein the pressure control valve is a solenoid control valve that controls a valve opening pressure of the valve body by adjusting a thrust of a plunger by an energization current to a coil. Adjustable shock absorber. The damping force adjusting type shock absorber according to claim 3, wherein a resistance force acting on the valve body is generated by a flow of fluid flowing through a gap between the valve body and the plunger. The fluid pressure buffer according to claim 3 or 4, wherein the valve body is pressed against the valve seat by the plunger via a valve spring interposed between the valve body and the plunger. vessel. 6. The fluid pressure shock absorber according to claim 4, wherein a resistance force acting on the valve body is increased by reducing a flow path area of the gap due to movement of the valve body in a valve opening direction. . The damping force adjusting buffer according to any one of claims 1 to 7, wherein the frequency of the predetermined high frequency input is an unsprung resonance frequency of a suspension device to which the damping force adjusting buffer is mounted. vessel. A cylinder filled with fluid, a piston slidably fitted in the cylinder, a piston rod having one end connected to the piston and the other end extending outside the cylinder, A main valve that controls the flow of fluid generated by sliding of the piston to generate a damping force; a back pressure chamber that controls the opening of the main valve by applying an internal pressure to the main valve in a valve closing direction; A fixed orifice for introducing fluid from the upstream side of the main valve to the back pressure chamber side, and a pressure control valve for controlling the flow of fluid from the back pressure chamber side to the downstream side of the main valve, In a damping force adjustment type shock absorber that controls the internal pressure of the back pressure chamber by a control valve,
The valve body of the pressure control valve has greater resistance to movement in the valve closing direction than resistance to movement in the valve opening direction, and a valve closing delay occurs with respect to a predetermined high frequency input. A shock absorber. 9. The damping force adjustment type according to claim 8, wherein the pressure control valve is a solenoid control valve that controls a valve opening pressure of the valve body by adjusting a thrust of a plunger by an energization current to a coil. Shock absorber. The damping force adjusting type shock absorber according to claim 9, wherein a resistance force acting on the valve body is generated by a flow of fluid flowing through a gap between the valve body and the plunger. The fluid pressure buffer according to claim 9 or 10, wherein the valve body is pressed against the valve seat by the plunger via a valve spring interposed between the valve body and the plunger. vessel. The fluid pressure buffer according to claim 10 or 11, wherein a resistance force acting on the valve body is increased by reducing a flow path area of the gap due to movement of the valve body in a valve opening direction. . The damping force adjustment type buffer according to any one of claims 8 to 12, wherein the predetermined high frequency input frequency is an unsprung resonance frequency of a suspension device to which the damping force adjustment type shock absorber is mounted. vessel. In a suspension control device in which a damping force adjustment type shock absorber is mounted between sprung springs of a suspension device of a vehicle, and the damping force of the damping force adjustment type shock absorber is adjusted by a controller based on the running state of the vehicle.
The damping force adjusting type shock absorber includes a cylinder filled with fluid, a piston slidably fitted in the cylinder, one end connected to the piston, and the other end extended to the outside of the cylinder. A piston rod, and a pressure control valve capable of generating a damping force by controlling a flow of fluid generated by sliding of the piston in the cylinder and adjusting a valve opening pressure in accordance with a control signal from the controller. Prepared,
The valve body of the pressure control valve has greater resistance to movement in the valve closing direction than resistance to movement in the valve opening direction, and a valve closing delay occurs with respect to a predetermined high frequency input. Suspension control device. The damping force adjusting buffer includes a main valve that generates a damping force by controlling a flow of fluid generated by sliding of the piston in the cylinder, and a back pressure that applies an internal pressure to the main valve in a valve closing direction. A part of the fluid flow is introduced into the back pressure chamber, the opening of the main valve is controlled by the internal pressure of the back pressure chamber, and the internal pressure of the back pressure chamber is controlled by the pressure control valve. The suspension control device according to claim 14, wherein the suspension control device is adjusted. 16. The suspension control according to claim 14, wherein the pressure control valve is a solenoid control valve that controls a valve opening pressure of the valve body by adjusting a thrust of a plunger by an energization current to a coil. apparatus. The suspension according to claim 16, wherein the valve body of the pressure control valve generates a resistance force acting on the valve body by a flow of fluid flowing through a gap between the valve body and the plunger. Control device. 18. The suspension according to claim 17, wherein the valve body of the pressure control valve is pressed against the valve seat by the plunger via a valve spring interposed between the valve body and the plunger. Control device. 18. The damping force adjusting shock absorber is characterized in that a resistance force acting on the valve body is increased by reducing a flow path area of the gap due to movement of the valve body in a valve opening direction. Or the suspension control apparatus of 18. 18. The suspension control device according to claim 14, wherein the predetermined high-frequency input frequency is an unsprung resonance frequency of the suspension device.

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