JPH04365980A - Hydraulic controller for vehicle - Google Patents

Abstract

(57) [Summary] This bulletin contains application data before electronic filing, so abstract data is not recorded.

Description

[Detailed description of the invention]

0001

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a hydraulic pressure control device suitable for use as a hydraulic pressure control source in an active suspension control system or a power steering system mounted on a vehicle.

0002

[Prior Art] Conventionally, in this type of hydraulic control device,
For example, as shown in Japanese Unexamined Patent Publication No. 52-79438, a constant discharge type pump is adopted as a hydraulic rotating machine, and when the load is not operating, the constant discharge type pump discharges the discharge according to its operation. Some pumps are designed to reduce the power loss of a constant discharge pump by returning oil to a drain tank by the unloading action of an unloading valve.

0003

[Problems to be Solved by the Invention] By the way, in such a configuration, since the above-mentioned pump is of a constant discharge type,
During the unloading action of the unload valve, the hydraulic pressure discharged from the pump changes suddenly, and as a result, the pump undergoes large rotational fluctuations due to the sudden change in its driving force. This means that the engine load fluctuates significantly, leading to a worsening of the driver's driving sensation. To deal with this, for example, as shown in Japanese Patent Application Laid-Open No. 60-187779, a variable displacement pump is adopted as the hydraulic rotary machine, and when the unloading valve is in the unloading operation, the variable displacement pump is used. It is also conceivable to reduce the capacity to suppress engine load fluctuations as described above. However, in such a case, the oil pressure to the load also decreases during the unloading action of the unload valve, and therefore it is not possible to reduce the pump capacity while maintaining the oil pressure to the load at a high level. . As a result, when the unloading valve performs the unloading action, the oil pressure applied to the load also unnecessarily decreases. Therefore, in order to cope with the above problems, the present invention employs a variable capacity hydraulic rotation means in a vehicle hydraulic pressure control device, and when the unload valve means unloads, the hydraulic pressure is applied to the load. The aim is to reduce the capacity of the hydraulic rotation means while maintaining the hydraulic rotation means appropriately.

0004

[Means for Solving the Problems] In order to solve the problems, the configuration of the present invention is such that the system is rotated by a prime mover mounted on a vehicle, sucks in hydraulic fluid in a reservoir tank according to its capacity, and discharges it to a load as pressurized fluid. a variable capacity hydraulic rotation means for discharging pressure fluid from the hydraulic rotation means to the load; and a check valve means for allowing only the discharge of pressure fluid from the hydraulic rotation means to the load; A main throttling means that throttles at a predetermined low throttling degree, and a main throttling means that directs the pressure fluid from the main throttling means to the reservoir in order to bring the hydraulic rotation means into an unloaded state when the hydraulic pressure on the downstream side of the main throttling means rises appropriately. an unload valve means for releasing the fluid into the tank and maintaining this released state in accordance with the hydraulic pressure on the downstream side of the check valve means; When the differential pressure between the upstream hydraulic pressure of the main throttle means and the downstream hydraulic pressure of the auxiliary throttle means is small, the downstream pressure fluid of the check valve means is throttled by the hydraulic rotation means. capacity control means for applying pressure fluid downstream of the check valve means to the hydraulic rotation means and controlling the fluid pressure to increase its capacity, and when the differential pressure is large, applying pressure fluid downstream of the check valve means to the hydraulic rotation means to decrease its capacity; The reason lies in the fact that it has been prepared for.

[0005]

[Operation] By configuring the present invention in this way, when the hydraulic rotation means is rotated by the prime mover,
If the hydraulic pressure to the load is low, the hydraulic rotation means sucks in the hydraulic fluid in the reservoir tank and discharges it to the load as pressurized fluid through the main throttling means and the check valve means. At this time, when the flow rate of the pressure fluid is small and the differential pressure between the upstream hydraulic pressure of the main throttle means and the downstream hydraulic pressure of the auxiliary throttle means is small, the capacity control means controls the check valve means. The downstream pressure liquid is applied to the hydraulic rotation means to increase its capacity. Thereafter, when the downstream hydraulic pressure of the main throttling means rises appropriately, the unload valve means directs the pressure fluid from the main throttling means into the reservoir in order to unload the hydraulic rotation means. The auxiliary throttling means is passed through the tank to release it, and this released state is maintained in accordance with the hydraulic pressure on the downstream side of the check valve means. Further, in such a stage, since the differential pressure increases, the capacity control means applies pressure fluid downstream of the check valve means to the hydraulic rotation means to reduce its capacity. Control.

[0006]

As described above, when the hydraulic rotation means is unloaded, the pressure fluid on the downstream side of the main throttle means is transferred to the auxiliary throttle while the hydraulic pressure to the load is properly maintained by the check valve means. The fluid is released into the reservoir tank through the means and the unload valve means, and this release is performed by reducing the capacity of the hydraulic rotating means, so load fluctuations and power loss of the hydraulic rotating means are minimized. At the same time, it becomes possible to unload the hydraulic rotation means.

[0007]

DESCRIPTION OF THE PREFERRED EMBODIMENTS A first embodiment of the present invention will be described below with reference to the drawings. FIGS. 1 and 2 show an example of a vehicle hydraulic pressure control device according to the present invention. This hydraulic pressure control device has a variable displacement hydraulic rotary machine P driven by the engine of the vehicle, and in this embodiment, a swash plate pump is adopted as the hydraulic rotary machine P. . The pump P has a pump housing 10, and the pump housing 10 has a sealing member 1 at an opening of a cylindrical housing member 10a.
It is constructed by coaxially assembling a disc-shaped housing member 10b via 0c. The rotating shaft 20 is coaxially supported in the pump housing 10 through both bearings 20a and 20b so as to be rotatable and immovable in the axial direction, and the outer end 21 of the rotating shaft 20 is connected to the output shaft of the engine. operably connected to. Note that in FIG. 2, reference numeral 22 indicates an oil seal.

A pair of cylinders 30 and 40 are disposed above and below the rotating shaft 20 as shown in FIG. 2, extending from the housing member 10b into the housing 10.
A piston 31 is fitted into the cylinder 40 so as to be slidable in the axial direction via a coil spring 32, while a piston 41 is fitted into the cylinder 40 so as to be slidable in the axial direction via a coil spring 42. ing. Further, the inside of the cylinder 30 communicates with the inside of the pipe P1 through the inflow port 11 of the housing member 10b, while the inside of the cylinder 40 communicates with the inside of the pipe P2 through the inflow port 12 of the housing member 10b. ing.

The swash plate 50 has its center hole 51 connected to the rotating shaft 2.
0, and its hemispherical base 52 is rotatably received by the spherical receiving surface 53a of the holder 53 (fitted in the bottom recess of the housing member 10a). On the upper end connecting portion 50a of the swash plate 50, a shoe 54 having a substantially spherical concave portion and connected to the head 31a of the piston 31 so as to be relatively rotatable but not able to escape is attached to the upper end connecting portion 50a.
It is installed so that it can move relative to 0a. On the other hand, the swash plate 50
A shoe 55, which has a substantially spherical concave portion and is connected to the head 41a of the piston 41 so as to be relatively rotatable but not detachable, is installed on the lower end connecting portion 50b so as to be movable relative to the lower end connecting portion 50b. Therefore, when the piston 31 moves to the right in FIG. 2 as the oil pressure inside the cylinder 30 increases, the swash plate 50
The piston 41 moves to the left against the coil spring 42 while tilting to the right. On the other hand, when the piston 41 moves to the right in FIG. 2 as the oil pressure within the cylinder 40 increases, the swash plate 50 tilts to the left while moving the piston 31 to the left against the coil spring 32. In such a case, the swash plate 5
0 to the right (or left) in FIG. 2 increases (or decreases) the capacity of the pump P.

[0010] The cylinder block 60 has both cylinders 30
, 40, and the rotating shaft 2 is located in the central hollow part 61 thereof.
0, and this cylinder block 6
In order to rotate integrally with the rotating shaft 20 of the cylinder block 60, the central hollow part 61 of the cylinder block 60 is splined to the left side of the rotating shaft 20. Further, between the central hollow part 61 of the cylinder block 60 and the annular member 62 (fitted to the middle part of the rotating shaft 20 so as not to be movable), a coil spring 63 is attached to the rotating shaft 20. The coil spring 63 is loosely fitted, and the coil spring 63 can slide on the bottom wall of the cylinder block 60 through the central hollow part 61 of the cylinder block 60 on the annular valve plate 70 (fixed to the inner surface of the housing member 10b). is in contact with.

A plurality of cylinders 6 of the cylinder block 60
Each cylindrical plunger 65, . . . 4, .
, 65 are fitted so as to be slidable in the axial direction, and the inside of each plunger 65 and the inside of each cylinder 64, which are located either forward or backward in the plane of the drawing with the rotating shaft 20 as the central axis, are Bottom wall opening 64a of cylinder 64
, the valve plate 70, a suction communication passage (not shown) of the housing member 10b, and a pipe P3 to communicate with the inside of the reservoir tank 80. On the other hand, the inside of each plunger 65 and the inside of each cylinder 64, which are located at a position symmetrical to the suction communication passage with the rotating shaft 20 as the central axis, are located in the bottom wall opening 64 of each cylinder 64.
a, communicates with the inside of the pipe P4 via a discharge passage (not shown) of the valve plate 70 and the housing member 10b. Each plunger 65
The spherical heads 65a are rotatably and irremovably received in substantially hemispherical recesses 66a of each shoe 66 that is slidably seated on the sliding surface 56 of the swash plate 50.

[0012] With this, each plunger 65 is connected to the swash plate 5.
Each shoe 66 slides in the axial direction within each cylinder 64 by each shoe 66 that slides in the circumferential direction on its sliding surface 56 in response to the tilting movement of 0.0. In other words, depending on the rotation of the rotating shaft 20 and the inclination of the swash plate 50, the hydraulic oil in the reservoir tank 80 is transferred to the rotating shaft 20.
As each plunger 65 in each cylinder 64 moves to the right during the upward displacement process of each cylinder 64 located below, piping P3, the suction communication passage of the housing member 10b, and the valve Each cylinder 64 sucks in hydraulic oil through the plate 70 and in this way.
With the displacement from above to below, the hydraulic oil in each cylinder 64 becomes pressure oil due to the leftward movement of each plunger 65, and flows through the valve plate 70 and the discharge communication passage of the housing member 10b to the pipe P4. discharged inside.

The hydraulic control mechanism W has a housing 90, within which a communication passage 91 is formed. The upstream portion 91a of this communicating path 91 communicates with the pipe P4, while the downstream portion 91b of this communicating path 91 communicates with the piping P5 connected to the accumulator 100 and the load. Upstream portion 91a of communication path 91
A main throttle 110 is interposed between the pump P and the midstream section 91c, and the main throttle 110 controls the discharge pressure oil flowing from the pump P through the pipe P4 into the communication passage 91 according to its predetermined throttle area. squeeze. This means that a pressure difference is generated between the upstream and downstream parts of the main throttle 110 depending on the throttle of the main throttle 110 and the flow rate. Note that in FIG. 1, the symbol 110a is
A circlip for fixing the main diaphragm 100 is shown.

The check valve 120 is provided in the housing 90 between the midstream section 91c and the downstream section 91b of the communication passage 91, so that the pressure oil flowing from the main throttle 110 into the midstream section 91c is directed into the downstream section 91b. It is interposed to allow only the inflow of water. This check valve 120 has a stepped cylindrical plug 121, and this plug 121 has a stepped inner part bored between the midstream part 91c and the downstream part 91b of the communication passage 91 in the housing 90. It is fitted coaxially within the hole 92. In such case,
The plug 121 has its male threaded portion 121a fastened to the female threaded portion 92a of the stepped inner hole 92.
An O-ring 122 is interposed between the large diameter portion 121b of the stepped inner hole 92 and the small diameter portion 92b of the stepped inner hole 92.

The cylindrical member 123 is fitted coaxially within the small diameter portion 92b and large diameter bottom portion 92c of the stepped inner hole 92, and the upper end of this cylindrical member 123 is fitted into the large diameter portion 121b of the plug 121. On the other hand, the bottom wall of the cylindrical member 123 is in contact with the bottom wall of the small diameter portion 92b via an O-ring 124. An inflow hole 123a is formed in the center of the bottom wall of the cylindrical member 123 so as to open into the midstream portion 91c of the communication passage 91. Further, at the lower part of the peripheral wall of the cylindrical member 123,
A plurality of outflow holes 123b, . , the downstream portion 91b of the communication path 91, and the large diameter bottom portion 9
It opens into a branch path 93 that branches off from 2c.

The stepped cylindrical valve body 125 is connected to the cylindrical member 123.
A valve body portion 1 is fitted coaxially and slidably inside the valve body 125 and is formed in a tapered shape on the outer circumference of the bottom wall of the valve body 125.
25a is coaxially seated on the annular inner end (which serves as a seating portion) of the inflow hole 123a of the cylindrical member 123. Further, a plurality of communication holes 125b, .
b and the communication path 91 through the annular passage of the large diameter bottom portion 92c.
It communicates with the downstream portion 91b of the branch path 91b and the branch path 93.

The coil spring 126 is loosely fitted coaxially into the small diameter portion of the plug 121 within the large diameter portion of the valve body 125. At the same time as locking the
The lower end thereof is engaged with a stepped portion of the inner circumferential surface of the valve body 125, and the valve body 125 is biased toward the seating portion of the cylindrical member 123. Therefore, in the check valve 120 configured as described above, when the resultant force of the elastic force of the coil spring 126 and the force due to the hydraulic pressure in the downstream portion 91b is larger than the force due to the hydraulic pressure in the midstream portion 91c of the communication passage 91. , the valve body 125 is the valve body part 1
At 25a, the cylindrical member 123 is seated on the seating part of the inflow hole 123a under the action of the coil spring 126, and the check valve 120 is closed, so that the downstream part 91b of the pressure oil in the midstream part 91c is closed.
Prevent the flow into the interior. On the other hand, the coil spring 126
When the resultant force of the elastic force and the force due to the hydraulic pressure in the downstream part 91b is smaller than the force due to the hydraulic pressure in the midstream part 91c, the valve body
The valve body portion 125a of No. 5 resists the coil spring 126 and dissociates from the seating portion of the cylindrical member 123 to open the check valve 120 and allow the pressure oil in the midstream portion 91c to flow into the downstream portion 91b. .

This means that the pressure oil in the downstream section 91b is applied to the accumulator 100 or the load via the pipe P5. Furthermore, when the check valve 120 is closed, the pressure oil stored in the accumulator 100 is transferred not only to the load but also to the pipe P5, the downstream portion 91b of the communication passage 91, each outlet hole 123b of the cylindrical member 123, and the valve body 125. It flows into the branch path 93 through each communication hole 125b. The auxiliary throttle 130 is interposed within the housing 90 between an upstream portion 94a and a midstream portion 94b of a branch passage 94 branched from the midstream portion 91c of the communication passage 91. This auxiliary aperture 13
0 has a stepped cylindrical plug 131, and this plug 131 is inserted into a stepped inner hole 95 bored between an upstream portion 94a and a midstream portion 94b of the branch passage 94 in the housing 90. Fitted coaxially. In such a case, the plug 131
is fastened in the female threaded part 95a of the stepped inner hole 95 at its male threaded part 131a, and between the small diameter part 131b of this plug 131 and the small diameter part 95b of the stepped inner hole 95,
An O-ring 132 is interposed.

The small diameter portion 131b of the plug 131 has a stepped inner hole 131c, and the large diameter portion thereof is connected to the upstream portion 9 of the branch passage 94.
It is coaxially drilled so as to communicate with 4a,
Inside the large diameter portion of this stepped inner hole 131c is a throttle member 133.
are coaxially fitted and prevented from coming off by a circlip 134. Accordingly, the throttle member 133 passes from the midstream portion 91c of the communication path 91 to the stepped inner hole 1 of the plug 131 via the upstream portion 94a of the branch path 94.
The pressure oil flowing into the large diameter portion of 31c is throttled. This means that
This means that a pressure difference is generated between the upstream and downstream parts of the throttle member 133 depending on the throttle of the throttle member 133. However, the predetermined aperture area of the aperture member 133 is considerably narrower than the predetermined aperture area of the main aperture 110. In addition, the small diameter portion 1 of the plug 131
31b, a communication passage 131d is bored perpendicularly to the small diameter portion of the stepped inner hole 131c, and this communication passage 131d is
Pressure oil flowing out from the throttle member 133 is caused to flow into the midstream portion 94b of the branch path 94.

The directional switching valve 140 is interposed within the housing 90 between the midstream portion 94b and the downstream portion 94c of the branch passage 94. This directional control valve 140 has a cylindrical plug 141, and this plug 141 is arranged in the middle part 94b and the downstream part 94c of the branch passage 94 in the housing 90.
It is coaxially fitted into a large diameter portion 96a of a stepped inner hole 96 bored between the two. In such a case, the plug 141 has its male threaded portion fastened to the female threaded portion of the large diameter portion 96a of the stepped inner hole 96, and the plug 141 and the large diameter inner hole 96
An O-ring 142 is interposed between the large diameter portion 96a and the large diameter portion 96a.

Further, the directional control valve 140 is connected to the spool valve 1.
40a, and this spool valve 140a has a cylindrical valve housing 143 coaxially fitted within the small diameter portion 96b of the stepped inner hole 96, and is slidable in the axial direction within this valve housing 143. It is composed of a fitted spool 144. The valve housing 143 has an annular flange portion 143a formed at its left end connected to the large diameter portion 96 of the stepped inner hole 96.
The valve housing 143 is stopped from coming out by being engaged with the stepped portion from the a side, and the right end opening of the valve housing 143 opens into the upstream portion 91a of the communication passage 91.

The spool 144 has an annular flange portion 1
At 44a, a coil spring 145 assembled in a cylindrical opening 141a formed coaxially on the tip side of the plug 141 is biased toward the open end side of the valve housing 143, and this spool 144 flange portion 14
4a is in contact with the flange portion 143a of the valve housing 143 and is prevented from coming off. The spool 144 has three land portions 144b, 144c, and 144d. When the land portion 144b is in the position shown in FIG.
b, cut off from the annular passage between 144c. In addition, when the land portion 144b is in the illustrated position in FIG.
144c. In addition, land portion 1
44c is the valve housing 14 when in the position shown in FIG.
The annular inner wall groove 143d of No. 3 is connected to both land portions 144c, 14
4d to communicate with the annular passage.

On the other hand, when the spool 144 moves to the left against the coil spring 145 because the force due to the pressure difference between the upstream section 91a and the midstream section 91c of the communication passage 91 is greater than the elastic force of the coil spring 145, Land portion 144
b connects the inner wall groove 143b of the valve housing 143 only to the annular passage between the land parts 144b and 144c, and the land part 144c communicates the inner wall groove 143c of the valve housing 143 only to the annular passage between the land parts 144c and 144d. In addition, the land portion 144d blocks the inner wall groove 143d of the valve housing 143 from the annular passage between the land portions 144c and 144d.

However, each inner wall groove 143b, 143c, 14
3d indicates each communication passage 97a bored in the housing 90;
97b and 97c, respectively, through the peripheral wall of the housing 90. Further, the annular passage between both land portions 144b and 144c and the annular passage between both land portions 144c and 144d communicate with communication passages 97d and 97e bored in housing 90, respectively, through the peripheral wall of valve housing 143. There is. Further, the communication passage 97a communicates with the inside of the reservoir tank 80 via the piping P6, and the communication passage 97b communicates with the inside of the reservoir tank 80 via the piping P7.
The communication path 97c communicates with the accumulator 100 via the downstream portion of the pipe P5, and the communication passage 97c communicates with the inside of the reservoir tank 80 via the pipe P8. Further, the communication path 97d communicates with the inside of the pipe P1, and the communication path 97e communicates with the inside of the pipe P2. Note that the midstream portion 94b and the downstream portion 9 of the branch path 94
4c are communicated through the respective communication holes 141b and 141c bored in the peripheral wall of the opening 141a of the plug 141 and the opening 14.
This is done through the inside of 1a.

In the directional control valve 140 constructed as described above, when the spool 144 is in the position shown in FIG. Passage 97b, inner wall groove 143c of valve housing 143, both land portions 144
b, annular passage between 144c, communication passage 9 of housing 90
7d, passes through the pipe P1 and the inlet port 11 of the pump P, and flows into the cylinder 30. As a result, piston 3
1 moves to the right and tilts the swash plate 50 to the right in order to increase the capacity of the pump P. At this time, the pressure oil in the cylinder 40 flows through the inflow port 12, the pipe P2, and the housing 90.
It is released into the reservoir tank 80 through the communication passage 97e, the annular passage between both land portions 144c and 144d, the inner wall groove 143d of the valve housing 143, the communication passage 97c of the housing 90, and the piping P8.

On the other hand, when the spool 144 is in the leftward movement position shown in FIG.
Downstream part of piping P5, piping P7, communication passage 97b of housing 90, inner wall groove 143c of valve housing 143, annular passage between land parts 144c and 144d, housing 90
It flows into the cylinder 40 through the communication passage 97e, the pipe P2, and the inflow port 12 of the pump P. This causes the piston 41 to move to the right and tilt the swash plate 50 to the left in order to reduce the capacity of the pump P. At this time, the pressure oil in the cylinder 30 flows through the inflow port 11, the pipe P1, the communication path 97d of the housing 90, and both land portions 144b, 144.
annular passage between c and inner wall groove 143b of valve housing 143;
, and is released into the reservoir tank 80 through the communication path 97a of the housing 90 and the piping P6.

The unload valve 150 is interposed within the housing 90 between the branch passages 93 and 94 and the outlet hole 98. This unload valve 150 has a pair of plugs 1
51, 152, each of these plugs 151, 1
52 is coaxially fitted into a stepped inner hole 99 bored in the housing 90 from each open end thereof. In such a case, the plug 151 has its male threaded portion 151a inserted into the stepped inner hole 9.
9, and the cylindrical portion 151b is inserted into the large diameter portion 99b of the stepped inner hole 99.
On the other hand, the plug 152 has its male threaded portion 152a fastened into the female threaded portion 99c of the stepped inner hole 99, and its cylindrical portion 152b is fitted into the stepped inner hole 99.
9 through the O-ring 152c. Further, a plurality of communication holes 151d, . The interior of the cylindrical portion 151b is communicated with the downstream end of the branch passage 93 via an annular groove 99e formed in the upper portion of the circumferential surface.

The cylindrical member 153 is coaxially fitted into the large diameter portion 99b of the stepped inner hole 99 between the annular stepped portion and the cylindrical portion 151b via both O-rings 153a and 153b. A plurality of communication holes 153c are bored in the peripheral wall of the cylindrical member 153 in the radial direction. Therefore, each communication hole 153c connects the hollow portion 153d of the cylindrical member 153,
The large diameter portion 99b of the stepped inner hole 99 communicates with the downstream end of the branch passage 94 via an annular groove 99f formed in a lower portion of the inner peripheral surface thereof. The stepped cylindrical spool 154 has a large diameter portion 1
54a is loosely fitted into the cylindrical portion 151b of the plug 151, and its medium diameter portion 154b and small diameter portion 154c are coaxially fitted into the hollow portion 153d of the cylindrical member 153. However, the middle diameter part 154b of the spool 154 is slidable within the hollow part 153d of the cylindrical member 153, and the axial length of the middle diameter part 154b is the same as the upper end of the cylindrical member 99 shown in FIG. It is shorter than the distance between the upper end of each communication hole 153c.

The poppet type valve body 155 is coaxially assembled between the cylindrical member 153 and a coil spring 156 that is engaged with the inner end of the plug 152 within the small diameter portion 99d of the stepped inner hole 99. The valve body 155 has a valve body portion 155a formed in a shape that tapers upward from the base of the valve body 155.
The unload valve 1 is selectively seated on the lower end portion (hereinafter referred to as a seating portion) of the hollow portion 153d of the cylindrical member 153 on the inclined peripheral wall under the biasing action of the coil spring 156.
It is designed to open 50. Further, the cylindrical tip portion 155b coaxially extends from the valve body portion 155a, and the cylindrical member 1
It is loosely fitted into the lower part of the hollow part 153d of the spool 153, and engages with the small diameter part 154c of the spool 154 coaxially and detachably from below. However, the axial length of the tip portion 155b of the valve body 155 is such that when the large diameter portion 154 of the spool 154 abuts the upper end of the cylindrical member 153 at its lower end, the valve body portion 155a of the valve body 155 is It is selected to open the unload valve 150 by resisting and disengaging the coil spring 156 from the seating portion. Note that the small diameter portion 9 of the stepped inner hole 99
The inside of 9d communicates with the inside of the reservoir tank 80 via the outflow hole 98 and piping P9.

In the unloading valve 150 configured as described above, as the oil pressure in the branch passage 94, that is, the oil pressure in the hollow part 153d of the cylindrical member 153 increases, the pressure in the hollow part 153d of the cylindrical member 153 of the valve body 155 increases. When the hydraulic pressure acting on the pressure receiving area becomes larger than the repulsive force of the coil spring 156, the valve body 155 opens against the coil spring 156, and the pressure oil in the branch passage 94 is transferred to the cylindrical member 153.
It is released into the reservoir tank 80 through the communication passages 153c, the hollow portion 153d, the small diameter portion 99d of the stepped inner hole 99, the outflow hole 98, and the piping P9. In this state, when the pressure oil in the accumulator 100 passes through the downstream portion 91b of the communication path 91, the check valve 120, and the branch path 93 and flows into the cylindrical portion 151b of the plug 151 from each communication hole 151d, the spool 154 is moved downward in response to the oil pressure in the cylindrical portion 151b, and its small diameter portion 154c engages with the tip portion 155b of the valve body 155 to maintain the open state of the valve body 155. When the hydraulic pressure acting on the pressure-receiving area of spool 154, that is, the area of 154b, becomes smaller than the elastic force of coil spring 156, valve body 155 moves upward and closes.

In this embodiment configured as described above, since the oil pressure in the hollow portion 153d of the cylindrical member 153 of the unloading valve 150 is low, the valve body 155 is moved under the action of the coil spring 156. Cylindrical member 1 at section 155a
It is assumed that the user is seated in the seating section No. 53. At this time, the rotational speed N of the pump P driven by the engine of the vehicle
If is low, the amount of pressure oil discharged from the pump P into the pipe P4 is small. Therefore, since the difference in oil pressure between the upstream and downstream parts of the main throttle 110 is small, the oil pressure in the upstream part 91a of the communication passage 91 and the oil pressure in the cylindrical part 141a of the plug 141, which act on both ends of the spool 144 of the spool valve 140, respectively. The differential pressure between them is smaller than the resiliency of the coil spring 145. Therefore, the spool 144 of the spool valve 140 is biased by the coil spring 145 and maintained in the position shown in FIG.

In such a state, the pressure oil from the accumulator 100 flows into the downstream part of the pipe P5, the pipe P5.
7. Communication path 97b of housing 90, directional switching valve 140
, and flows into the cylinder 30 of the pump P via the inflow port 11 through the communication path 97d and the piping P1. At this time,
The pressure oil in the cylinder 40 is connected to the inflow port 12, piping P2,
Communication path 97e of housing 90, both land portions 144c,
Annular passage between 144d and inner wall groove 1 of valve housing 143
43d, communication path 97c of housing 90 and piping P8
It is released into the reservoir tank 80 through. Therefore, the capacity of the pump P reaches its maximum capacity (the state shown in FIG. 2) due to the sliding of the piston 31 protruding from the cylinder 30 and the sliding of the piston 41 into the cylinder 40 against the coil spring 42. increase towards

As the rotational speed N of the pump P increases, the amount of pressure oil discharged from the pump P into the pipe P4 increases, and the difference in oil pressure between the upstream and downstream parts of the main throttle 110 also increases. Then, the hydraulic pressure difference acting on both ends of the spool 144 is
When the elastic force increases beyond the elastic force of the coil spring 145, the spool 144 slides inside the valve housing 143 against the coil spring 145. When the spool 144 moves a certain amount to the left, the pressure oil in the accumulator 100 is transferred to the pipe P5.
downstream part, piping P7, communication passage 97b, directional switching valve 140
It flows into the cylinder 40 of the pump P via the inflow port 12 through the communication passage 97e and the pipe P2. At this time,
The pressure oil in the cylinder 30 is connected to the pipe P1, the communication path 97d,
It is released into the reservoir tank 80 through the directional switching valve 140, the communication path 97a, and the pipe P6. Therefore, the capacity of the pump P decreases toward the minimum capacity as the piston 41 slides out of the cylinder 40 and the piston 31 slides into the cylinder 30 against the coil spring 32.

After that, the above-described changes in the capacity of the pump P due to the fluctuations in the rotational speed N are repeated, and the hydraulic pressure difference between the upstream and downstream parts of the main throttle 110 is controlled to be constant. As a result, if the capacity of the pump P is α and the discharge pressure oil amount of the pump P is Q, then at a certain rotation speed or higher,
The discharge pressure oil amount Q remains constant regardless of fluctuations in the rotational speed N (see FIG. 3(A)). This means that Nα is constant. Under such a state, the check valve 120 is selected depending on the difference between the resultant force of the coil spring 126 and the hydraulic pressure in the downstream portion 91b of the communication path 91 and the force due to the hydraulic pressure in the midstream portion 91c. Open the pump P
The pressure oil discharged from the downstream part 91b of the communication passage 91 and the pipe P
5 to the accumulator 100 and the load. Note that in such a state, the unload valve 150
Since the is closed, no flow occurs in the branch passage 94, and therefore no flow of pressure oil occurs in the upstream and downstream portions of the auxiliary throttle 130. As a result, no oil pressure difference occurs between the upstream and downstream parts of the auxiliary throttle 130.

After that, when the oil pressure in the accumulator 100 rises to the set pressure, the pressure inside the hollow part 153d of the cylindrical member 153 of the unload valve 150 flows through the midstream part 91c of the communication passage 91, the branch passage 94, and the auxiliary throttle 130. The oil pressure also increases, causing the valve body 155 to move downward against the coil spring 156. Therefore, the pressure oil in the midstream portion 91c of the communication passage 91 is transferred to the hollow portion 153d of the cylindrical member 153 of the unload valve 150, the small diameter portion 99d of the stepped inner hole 99, the outflow hole 98, and the pipe P9.
and is released into the reservoir tank 80. This means that the pump P is placed in an unloaded state.

In addition, in this unloaded state, the pressure oil in the midstream section 94b and the downstream section 94c of the branch passage 94 is released into the reservoir tank 80 as described above, but the pressure oil in the upstream section of the branch passage 94 The oil pressure in the upstream part 94a of the branch passage 94 is maintained high due to the throttling action of the auxiliary throttle 130, and the oil pressure in the upstream part 91a of the communication passage 91 is maintained high due to the throttling action of the main throttle 110. It is maintained even higher than the oil pressure inside. Therefore, the spool 14 of the directional valve 140
4 resists the coil spring 145 due to the high oil pressure in the upstream portion 91a of the communication path 91 and slides to the left moving end. Therefore, the pressure oil in the accumulator 100 is transferred to the pipe P7.
, the communication passage 97b, the directional switching valve 140, the communication passage 97e, and the pipe P2, and flows into the cylinder 40 of the pump P via the inflow port 12. As a result, the capacity of the pump P becomes almost the minimum capacity, and the amount of pressure oil discharged from the same pump P also becomes the minimum amount.

Furthermore, when the pump P is in the unloaded state as described above, the hydraulic pressure in the upstream portion of the check valve 120 (that is, the midstream portion of the communication passage 91) flows through the branch passage 94, the auxiliary throttle 130, and the is released into the reservoir tank 80 via the unload valve 150, downstream of the check valve 120 (i.e.
The oil pressure in the downstream portion 91b) of the communication passage 91 is at a set pressure due to the oil pressure in the accumulator 100. Therefore, the valve body 125 of the check valve 120 is seated on the seat portion of the cylindrical member 123 with its valve body portion 125a, thereby blocking communication between the midstream portion 91c and the downstream portion 91b of the communication passage 91.

Thus, the pressure oil in the accumulator 100 passes through the downstream portion 91b of the communication path 91, the check valve 120, and the branch path 93, and enters the cylindrical portion 151b of the plug 151 of the unload valve 150 through the communication holes 151d. When it flows in from
Since the oil pressure in the cylindrical portion 151b is greater than the sum of the oil pressure in the hollow portion 153d of the cylindrical member 153 and the elastic force of the coil spring 156, the spool 154 is attached to the tip of the valve body 155 at its small diameter portion 154c. The portion 155b is pushed downward to maintain the open state of the valve body 155.

As explained above, when the pump P enters the unloaded state, the capacity of the pump P immediately becomes almost the minimum capacity, and the discharge pressure oil amount also becomes the minimum amount, so that the reservoir tank 80 It is possible to achieve an unloaded state of the pump P while minimizing the amount of pressure oil discharged from the pump P to be recirculated. Therefore, the power loss of the pump P can be reduced to a minimum. In this case, the amount of pressure oil returned to the reservoir tank 80 during unloading is suppressed to the minimum as described above, so the opening area and shape of the unload valve 150 are reduced. This means that the unload valve 150 can be made more compact.

Furthermore, when the pump P enters the unloaded state, the capacity of the pump P is immediately maintained at the minimum capacity, so that fluctuations in the driven force caused by the unloading of the pump P are alleviated. Therefore, fluctuations in engine speed are suppressed, which helps improve the driving sensation of the vehicle. In addition, as described above, during unloading, the amount of pressure oil discharged from the pump P is the minimum amount, so the main throttle 110 with a low degree of throttling
Almost no oil pressure difference occurs between the upstream and downstream parts. Therefore, the sliding position of the spool 144 of the directional control valve 140 is specified by the hydraulic pressure difference between the upstream and downstream parts of the auxiliary throttle 130, which has a high degree of restriction, and the capacity of the pump P is determined by the hydraulic pressure difference between the upstream and downstream parts of the auxiliary throttle 130. It is controlled around the minimum capacity to keep the difference constant. As a result, even if the rotation speed of the pump P changes, the discharge oil pressure of the pump P does not change, and the power loss of the pump P is reduced.

Thereafter, when the oil pressure in the accumulator 100 gradually decreases from the set pressure in accordance with consumption due to the load, the cylindrical portion 151b of the plug 151 of the unload valve 150
The oil pressure inside will also drop. Therefore, the valve body 155
It is gradually pushed up by the elastic force of the coil spring 156, and the valve body part 155a is seated on the seating part of the cylindrical member 153, and the hollow part 153d of the cylindrical member 153 is inserted into the large diameter inner hole 99.
It is cut off from communication with the small diameter portion 99d. Then, the hydraulic pressures in the intermediate portions 91c of the branch passage 94 and the communication passage 91 rise, and the check valve 120 opens. Then, a flow of pressure oil occurs between the upstream and downstream parts of the main throttle 110, and the spool 1 of the directional control valve 140
44 slides so as to keep the differential pressure between the upstream and downstream parts of the main throttle 110 constant, and as the capacity of the pump P increases, the discharge pressure oil amount of the pump P also returns.

In such a case, if a certain area difference is provided between the pressure receiving area of the spool 154 of the unload valve 150 and the pressure receiving area of the valve body 155, the hydraulic pressure for opening and closing the valve body 155 (in other words, , the set pressure for unloading, and the return pressure from unloading) can have hysteresis (see FIG. 3(B)). In addition, in FIG. 3(B), the symbol x indicates the accumulator pressure corresponding to the set pressure for unloading, and the symbol y indicates the accumulator pressure corresponding to the return pressure from unloading. Further, the symbol H indicates the width of the above-mentioned hysteresis. Further, the set pressure for unloading can be changed by adjusting the elastic force of the coil spring 156 by adjusting the position of the plug 152.

Next, a second embodiment of the present invention will be described with reference to FIGS. 4 and 5.
To explain with reference to the above, in this embodiment, the first
In place of the pump P and directional switching valve 140 described in the embodiment, a pump Pa and directional switching valve 140A are adopted, and each communication path 97c is connected from the housing 90 in the first embodiment.
, 97d and the respective pipes P1 and P8 are omitted. Pump P
In the pump P described in the first embodiment, a swash plate 52A is used instead of the swash plate 52, and a spring mechanism 30A is used instead of the inlet port 11, cylinder 30, piston 31, and coil spring 32. This is what was adopted. The spring mechanism 30A is interposed between the lower end of the swash plate 52A shown in FIG. 5 and the bottom of the housing member 10a so as to coaxially face the cylinder 40 via the lower end of the swash plate 52A. . This spring mechanism 30A
has a coil spring 33, one end of which is engaged with the bottom of the housing member 10a, and the other end of which is engaged with the lower end of the swash plate 52A via a receiving member 34. It is assembled. In such a case, the receiving member 34 is attached to the swash plate 5 with its spherical head 34a.
It is rotatably engaged in a spherical recess formed at the lower end of 2A.

The directional control valve 140A is the same as the directional control valve 140 described in the first embodiment.
3 and the spool 144 are replaced with a valve housing 146 and a spool 147. Spool 147
has three land portions 147a, 147b, and 147c, and the land portion 147a connects the annular inner wall groove 146a of the valve housing 146 to the annular passage between the land portions 147a and 147b when in the position shown in FIG. communicate. Further, when the land portion 147b is in the illustrated position in FIG. 4, the annular inner wall groove 146b of the valve housing 146
7a and 147b to communicate with the annular passage. Furthermore, when the land portion 147c is in the illustrated position in FIG.
147c. On the other hand, when the spool 147 slides to the left moving end in the same manner as the sliding of the spool 144 to the left moving end in the first embodiment, the land portion 147b moves the annular inner wall groove 146b between both land portions 147b,
147c is communicated with the annular inner wall groove 146c via an annular passage. However, each annular inner wall groove 146a, 146b, 14
6c is connected to each communication passage 97 through the peripheral wall of the valve housing 146.
a, 97e, and 97b, respectively. Note that the spring mechanism 30A serves as a capacity increasing mechanism substantially similar to the cylinder 30 and piston 31 in the first embodiment. The rest of the structure is the same as that of the first embodiment.

In the second embodiment configured as described above, the cylindrical portion 141a of the plug 141 of the directional control valve 140A is
Since the differential pressure between the oil pressure inside the coil spring 1 and the oil pressure inside the upstream portion 91a of the communication path 91 is small, the spool 147
45, the pressure oil in the cylinder 40 of the pump Pa flows through the inlet ports 12,
It is released into the reservoir tank 80 through the pipe P2, the communication passage 97e of the housing 90, the directional switching valve 140A, the communication passage 97a, and the pipe P6. Therefore, the spring mechanism 30A pushes the lower end of the swash plate 52A against the coil spring 42 by the elastic force of the coil spring 33, and tilts the swash plate 52A to an angle corresponding to the maximum capacity of the pump Pa.

On the other hand, the plug 141 of the directional control valve 140A
The hydraulic pressure in the cylindrical part 141a and the upstream part 91a of the communication path 91
When the spool 147 slides to the left against the coil spring 145 when the difference between the oil pressure in the communication passage 9 and the
7e is blocked from the communication path 97a by the land portion 147b, and communicated with the communication path 97b via the annular path between the land portions 147b and 147c. Therefore, the pressure oil in the accumulator 100 is transmitted to the downstream part of the pipe P5, the pipe P7, the communication passage 97b, the directional control valve 140A, and the communication passage 9.
7e and piping P2 into the cylinder 40 of the pump Pa through the inflow port 12, the piston 41 slides and protrudes from the cylinder 40 against the coil spring 33 of the spring mechanism 30A and pushes the lower end of the swash plate 52A to the right. move towards. Therefore, the capacity of pump Pa decreases toward the minimum capacity. This can also be achieved by adopting a spring mechanism 30A instead of the piston 31 and cylinder 30.
This means that the same effects as in the first embodiment can be achieved.

In the second embodiment, the spring mechanism 30A is disposed coaxially with the piston 41, but instead of this, the cylinder 30 and the piston 41 described in the first embodiment are arranged. A spring mechanism 30A may be provided at the location where the piston 31 is provided. In addition, instead of the pressure oil from the accumulator 100, pressure oil on the upstream side or downstream side of the main throttle 110 is passed through the pipe P7 to the communication passage 97.
It may be added within b.

Next, a first modification of the first or second embodiment will be explained with reference to FIG. 6. In this first modification, the features described in the first or second embodiment are The structure is characterized by the adoption of an unload valve 150A in place of the unload valve 150. Unload valve 150A
In the above-described unload valve 150, the cylindrical member 15
3. In place of the spool 154 and the valve body 155, a cylindrical member 153A, a stepped cylindrical spool 156, and a rod-shaped valve body 15
7 was adopted.

The cylindrical member 153A, like the cylindrical member 153, is fitted into the large diameter part 99b of the stepped inner hole 99 of the housing 90, and inside this cylindrical member 153A, there is a large diameter hollow part 153e and a small diameter hollow part 153e. Hollow portions 153f are formed coaxially above and below. A pair of communication passages 153g and 153h are bored in the radial direction on one side of the peripheral wall of the cylindrical member 153A, and the communication passage 153g opens at its inner end into the lower part of the large diameter hollow portion 153e. However, its outer end communicates with the downstream portion 94c of the branch path 94. Further, the communication passage 153h has a small diameter hollow portion 153 at its inner end.
The outer end of this communication path 153h communicates with an additional branch path 94d that additionally branches from the downstream end of the downstream portion 94c of the branch path 94. Further, an L-shaped communication passage 153i is bored on the other side of the peripheral wall of the cylindrical member 153A, and the communication passage 53i has a small diameter hollow at its upper end so as to face the inner end of the communication passage 153h. Part 153
f, and its lower end opens into the small diameter portion 99d of the stepped inner hole 99.

The spool 156 is loosely fitted into the cylindrical portion 151b of the plug 151 at its large diameter portion 156a, and the medium diameter portion 156b of the spool 156 is loosely fitted into the cylindrical portion 151b of the plug 151.
It is coaxially and slidably fitted into the large diameter hollow part 153e of 3A. Further, the small diameter portion 156c of the spool 156 is coaxially and loosely fitted into the small diameter hollow portion 153f of the cylindrical member 153A. The rod-shaped valve body 157 is fitted coaxially and slidably from below into the small-diameter hollow portion 153f of the cylindrical member 153A, and is located at an intermediate portion between the base end 157a and the distal end 157b of the valve body 157. A small diameter portion 157c is formed to provide an annular communication path that selectively communicates with both communication paths 153h and 153i. Further, the valve body 157 is biased upward in FIG. 6 by a coil spring 156 at its receiving portion 157d. However, in this modification, when the receiving portion 157d of the valve body 157 comes into contact with the lower end surface of the cylindrical member 153A, the base end portion 157a connects the two communicating passages 153h, 153i.
The axial length of the base end portion 157a is determined such that the base end portion 157a is cut off from communication with the small diameter hollow portion 153f. Also, spool 1
When the large diameter portion 156a of the valve body 157 comes into contact with the upper end surface of the cylindrical member 153A, the annular communicating passage 1 of the valve body 157
53h, 153i so as to have sufficient communication with spool 1.
56 medium diameter portion 156b, small diameter portion 156c and valve body 157
The sum of the lengths of each axis of the tip portion 157b is determined. The other configurations are the same as those of the first or second embodiment.

In the first modified example configured in this way, the cylindrical member 153 communicating with the downstream portion 94c of the branch path 94
As the oil pressure in the large diameter hollow part 153e of A increases, the valve body 1
When the hydraulic pressure acting on the pressure-receiving area of the tip 157b of 57 becomes larger than the elastic force of the coil spring 156, the valve body 157 slides downward against the coil spring 156, causing the tip 157b and the base 157a is communicated with both communication paths 153h and 153i. Then, the pressure oil in the branch passage 94 is transferred to the additional branch passage 94d and the communication passage 1.
53h, the small diameter hollow part 153f, the communication passage 153i, the small diameter part 99d of the stepped inner hole 99, the outflow hole 98, and the pipe P9, and are released into the reservoir tank 80. Further, in this case, the cylindrical portion 15 of the plug 151 that communicates with the inside of the branch path 93
Pressure oil flows into 1b from the accumulator 100, the spool 156 moves downward, and the small diameter portion 156c closes the valve body 157.
The valve body 157 is engaged with the distal end portion 157b of the valve body 157 to maintain the downwardly moving state of the valve body 157 as it is. This also applies to the first or second valve depending on the unloading action of the unloading valve 150A.
This means that the same effects as in the embodiment can be achieved. Note that other effects are the same as those of the first or second embodiment.

FIG. 7 shows a partial modification of the first modification. In this partial modification, instead of the cylindrical member 153A described in the first modification, the cylindrical member 15
Its structural features lie in the adoption of 3B. The cylindrical member 153B is connected to the housing 9 similarly to the cylindrical member 153A.
This cylindrical member 1 is fitted into the large diameter portion 99b of
A communication passage 153j is bored in the radial direction on one side of the peripheral wall of 53B. Further, this communication passage 153j communicates with the downstream portion 94c of the branch passage 94 at its outer end, and the peripheral wall opening of this communication passage 153j is connected to the L-shaped communication passage 153.
It communicates with the lower part of the large-diameter hollow part 153e through k. The communication passage 153j opens into the small diameter hollow part 153f at the inner end opening 153l, and the cross-sectional shape of the inner end opening 153⌒ is isosceles as shown in FIG. 8(A). It is triangular. In this case, the apex and bottom of this isosceles triangle correspond to the upper and lower ends of the cross-sectional shape of the inner end opening 153l shown in FIG. 7, respectively. The other configurations are the same as the first modification.

In this partial modification configured in this way, the valve body 157 is caused to close to the coil due to the increase in hydraulic pressure generated in the large diameter hollow portion 153e of the inner cylinder member 153B through the branch passage 94 and both communicating passages 153j and 153k. When sliding downward against the spring 156 in the same manner as in the first modification, the annular passage between the proximal end 157a and the distal end 157b of the valve body 157 connects with the inner end opening 153l of the communication passage 153j. communicate. In such a case, since the inner end opening 153l has an isosceles triangular cross section as described above, the communication cross-sectional area of the annular passage with the inner end opening 153l extends downward from the valve body 157. It gradually increases as the surface slides. In the process of increasing the communication cross-sectional area, at the beginning, the communication cross-sectional area is narrower than the opening area of the auxiliary aperture 130, so the auxiliary aperture 1
A state in which back pressure is applied to 30 occurs. Therefore, after the unloading action of the unloading valve 150 starts, the spool 14
As shown in FIG. 8(D), each oil pressure acting on both ends of the cylinder 4 or 147 gradually decreases over time. As a result, the spool 144 or 147 does not move back and forth due to large fluctuations in the oil pressure between both ends of the pump P or P during unloading.
The capacitance of a can be smoothly changed.

In the above partial modification, the inner end opening 153l of the communication path 153j has an isosceles triangular cross-sectional shape; however, the inner end opening 153 is not limited to this.
Substantially the same effect as the above partial modification example can be obtained even if the cross-sectional shape of l is as shown in FIG. 8(B) or FIG. 8(C).

FIG. 9(A) shows a second modification of the first or second embodiment, and in this second modification, the features described in the first or second embodiment are Its structural feature lies in the fact that a linear actuator 110A is used instead of the main diaphragm 110. Linear actuator 11
0A attracts the rod-shaped valve body 112 against the coil spring 113 according to the attraction force based on the inflow current I of the solenoid 111, and connects the upstream portion 91a and the midstream portion 91 of the communication path 91.
The degree of communication with c (ie, the degree of aperture) is controlled. In such a case, if the inflow current I is increased to I1, the upper midstream portion 91
The degree of restriction between a and 91c is reduced, increasing the flow rate of pressurized oil (see FIG. 9(B)). On the other hand, when it decreases to I=I2 (<I1), the degree of throttling between the upper and midstream portions 91a and 91c increases and the pressure oil flow rate decreases (see FIG. 9(B)). This means that the discharge pressure oil amount can be set arbitrarily by adjusting the inflow current I. Furthermore, the linear actuator 110A may be employed as the auxiliary throttle 130, and in such a case, the amount of pressure oil returned to the reservoir tank 80 when the unload valve 150 is unloaded can be set arbitrarily.

FIG. 10 shows a third modification of the second embodiment, in which the relief valve 1
60 is an upstream portion 94 of the branch passage 94 within the housing 90
A bypass 94 formed to extend from the intermediate portion of a.
Its structural feature lies in that it is interposed between the bypass 94f formed to communicate with the communication passage 131d of the auxiliary throttle 130 and the auxiliary throttle 130. Therefore, in the relief valve 160, by setting the elastic force of the coil spring 161, when the pump Pa is unloaded, the upstream portion 94 of the branch passage 94 is
The oil pressure in a can be freely set via the bypass 94e. Therefore, even if the discharge pressure oil amount increases due to an increase in the rotational speed of the pump Pa during unloading, the pressure oil generated in the upstream portion 94a of the branch passage 94 is transferred to the throttle member 133 of the auxiliary throttle 130 by the relief valve 160. If the air is allowed to flow directly into the midstream portion 94b via both bypasses 94e and 94f without flowing in, the increase in the differential pressure between the upstream and downstream portions of the auxiliary throttle 130 can be suppressed to a minimum. As a result, not only is it useful for saving power, but the spool 144 or 147 can be maintained at the pressure end, and the operation of the directional control valve 140A during unloading can be stabilized. Can be maintained. This holds true in the case of the directional control valve 140 as well. Note that the first
In the second embodiment, an unload valve 150 is used, but instead of this unload valve 150, as shown in FIG. An additional configuration of an unload valve may be employed.

Next, a third embodiment of the present invention will be described with reference to FIGS.
4, in this third embodiment,
In place of the main throttle 110 and unload valve 150 described in the first or second embodiment, a linear actuator 110B and a proportional electromagnetic relief valve 180 are adopted, respectively, as shown in FIGS. 12 and 13, A feature of the structure is that an electric control circuit E for driving and controlling the linear actuator 110B and the relief valve 180 is employed as shown in FIG. 14. The linear actuator 110B has the same configuration and function as the linear actuator 110A (see FIG. 9(A)) described in the second modification of the first or second embodiment,
Like the linear actuator 110A, this linear actuator 110B responds to the attraction of the rod-shaped valve body in accordance with the attraction force based on the current I flowing into the solenoid, and the communication path described in the first or second embodiment. Upstream part 9 of 91
It is interposed in the communication passage 91 so as to control the degree of communication (that is, the degree of constriction) between 1a and the midstream portion 91c.

The relief valve 180 is equipped with a poppet type valve body 181, as shown in FIG.
1 is the housing 9 described in the first or second embodiment.
In FIG. 1 or 4 of FIG. 0, the inner hole 99A is formed on the right side wall between the downstream portion 94c of the branch passage 94 and the pipe P9, and is coaxially displaceable. Thus, the valve body 181 receives the biasing force of a coil spring 182, which will be described later, at its conical valve body portion 181a, and the inflow hole of the inner hole 99A (which communicates with the downstream portion 94c of the branch passage 94). Seated at the inner end. Moreover, this valve body 181 is
In response to the displacement against the biasing force of the coil spring 182, the valve body portion 181a is separated from the inner end of the inflow hole of the inner hole 99A, and the downstream portion 94c of the branch passage 94 is communicated with the outlet hole of the inner hole 99A. It communicates with the inside of the pipe P9. However, the degree to which the downstream portion 94c of the branch passage 94 is communicated with the inside of the pipe P9 depends on the degree of dissociation of the inner hole 99A of the valve body portion 181a from the inner end of the inflow hole (corresponding to the amount of displacement of the valve body 181). Match.

The relief valve 180 also includes a linear actuator 180a, as shown in FIG.
It is assembled coaxially with the inner hole 99A on the right side wall portion of No. 0. The linear actuator 180a has a cylindrical casing 183, and the annular flange 183a of the casing 183 is arranged on the right side wall of the housing 90 so as to coaxially face the inner hole 99A. It is fixed to the part. The solenoid 184 is
Fitted concentrically within the casing 183, this solenoid 184 generates an attractive force proportional to its inflow current J. The outer circumferential surface of the disk 185 is fitted liquid-tightly and coaxially into the opening of the casing 183 via an O-ring 185a. It is in liquid-tight contact with the right side wall portion of the housing 90. Note that the disc 185 communicates the inside of the casing 184 with the inside of the inner hole 99A through its communication hole 185c.

The stepped rod 186 is made of a magnetic material, and includes a large diameter portion 186a coaxially and loosely fitted into the hollow portion of the solenoid 184, and a large diameter portion 186a coaxially fitted into the hollow portion of the solenoid 184.
86a and a small diameter portion 186b extending coaxially from 86a. The rod 186 has a circular hole-shaped receiving portion 1 formed in the lower wall of the casing 183 at the base of the large diameter portion 186a.
83b so as to be coaxially movable, and the small diameter portion 1
86b is rotatably supported within the shaft hole 185d of the disc 185 and is assembled concentrically within the casing 183. Thus, the rod 186 is attracted by the attraction force of the solenoid 184 and is displaced in the axial direction. The coil spring 182 described above is connected to the annular flange 181 of the valve body 181.
b and an annular flange 187a of a locking member 187 coaxially fitted to the tip of the small diameter portion 186b of the rod 186, concentrically with the inner hole 99A, and this coil spring 182 constantly urges the valve body 181 to the right in FIG. However, this coil spring 182
The set load is determined by the amount of displacement of the rod 186.

The electric control circuit E has a vehicle speed sensor 200, as shown in FIG.
0 detects the actual vehicle speed of the vehicle and generates a vehicle speed signal at a frequency proportional to this. The waveform shaper 210 shapes the waveform of the vehicle speed signal from the vehicle speed sensor 200 and generates a shaped signal. The steering angle sensor 220 detects the steering angle of the steering wheel of the vehicle and generates a steering angle detection signal. The acceleration sensors 230, 240, 250, and 260 are respectively disposed on the left front wheel, the right front wheel, the left rear wheel, and the right rear wheel of the vehicle. Accelerations G1 to G4 generated in the lateral direction during traveling are detected and generated as first to fourth lateral acceleration detection signals. The A-D converter 270 receives the steering angle detection signal from the steering angle sensor 220 and the first to fourth signals from each of the acceleration sensors 230 to 260.
The lateral acceleration detection signals are respectively converted into digital signals and generated as a steering angle digital signal and first to fourth lateral acceleration digital signals.

The microcomputer 280 executes the computer program in cooperation with the waveform shaper 210 and the A-D converter 270 according to the flowchart shown in FIG. Each drive circuit 290a, 2 is connected to the solenoid 184 of the valve 180, respectively.
Performs arithmetic processing necessary to drive and control 90b. however,
The above computer program is stored in the ROM of the microcomputer 280 in advance. Note that the microcomputer 280 operates by being supplied with power from the battery B via the ignition switch IG of the vehicle. The other configurations are the same as those of the first or second implementation order.

In the third embodiment configured as described above, if the vehicle is caused to run with its engine started in response to the closing of the ignition switch IG, and the microcomputer 280 is activated, the same result can be achieved. The microcomputer 280 executes the computer program in FIG.
Execution starts at step 300a according to the flowchart of No. 5. In step 310, the microcomputer 280 receives the value of the steering angle digital signal from the A-D converter 270 (hereinafter referred to as steering angle θ) and the first to
Each value of the fourth lateral acceleration digital (hereinafter, each lateral acceleration G
1 to G4).

[0064] Next, the microcomputer 280
In step 320, the maximum value of each of the lateral accelerations G1 to G4 in step 310 is selected, and the maximum lateral acceleration Gm
ax, and in the next step 330, the inflow current I
The inflow currents Ia and Ja are determined according to the maximum lateral acceleration Gmax in step 320 based on the characteristic curve La (see FIG. 16(B)) representing the relationship between a (or Ja) and the maximum lateral acceleration Gmax. However, the characteristic curve La is previously stored in the ROM of the microcomputer 280 as data. In such a case, each inflow current Ia and Ja represents a current component flowing into the solenoid of the linear actuator 110B and the solenoid of the relief valve 180, respectively, in relation to the maximum lateral acceleration Gmax.

After that, the microcomputer 280
In step 340, the actual vehicle speed of the vehicle (hereinafter referred to as vehicle speed V) is calculated according to the shaping signal from the waveform shaper 210, and the relationship between the vehicle speed V and the estimated current Ib (or Jb) is calculated. Each curve L represents the steering angle θ as a parameter.
b (see FIG. 16(A)) and the vehicle speed V and step 3.
Each estimated current Ib and Jb according to the steering angle θ at 10
Determine. However, each curve Lb is stored in advance in the ROM of the microcomputer 280 as data. In this case, each of the estimated currents Ib and Jb represents a current component that is estimated to flow into the solenoid of the linear actuator 110B and the solenoid of the relief valve 180 in relation to the vehicle speed V and the steering angle θ. Note that this method focuses on the fact that the lateral acceleration G corresponds uniquely to the estimated current Ib (or Jb) on the premise that the lateral acceleration G is proportional to the product of the vehicle speed V to the nth power and the steering angle θ. be.

[0066] The microcomputer 280 calculates the inflow current Ia at step 350 based on the following equation 1.
and calculate an inflow current I according to the estimated current Ib, and
The inflow current J is calculated according to the inflow current Ja and the estimated current Jb based on the following equation 2.

[Math. 1] I=CIa+Ib

[Equation 2] J=CJa+Jb However, in Equations 1 and 2, the symbol C represents a control gain. Further, Equation 1 and Equation 2 are stored in advance in the ROM of the microcomputer 280.

After that, the microcomputer 280
At step 360, each inflow current I and J is generated as a first and second output signal, respectively. Then, the drive circuit 290a causes the inflow current I to flow into the solenoid of the linear actuator 110B in response to the first output signal from the microcomputer 280, and causes the solenoid to generate an attractive force proportional to the inflow current I. , drive circuit 2
90b causes an inflow current J to flow into the solenoid of the linear actuator 180a in response to a second output signal from the microcomputer 280, and generates an attractive force proportional to the inflow current J from the solenoid. Therefore, the linear actuator 110B sucks the valve body according to the suction force from the solenoid and controls the degree of constriction between the upstream section 91a and the midstream section 91c of the communication path 91, while the linear actuator 180a The rod 186 is attracted according to the attraction force of the coil spring 182, and the set load of the coil spring 182 is set and controlled.

In such an operating process, the hydraulic rotating machine and the hydraulic control mechanism (
When the main throttle 110 and the unload valve 150 (excluding the main throttle 110 and the unload valve 150) are operating in substantially the same manner as described in the first or second embodiment, the linear actuator 110B is activated in response to the suction of the valve body. Upstream section 91a and midstream section 9 of passage 91
1c is controlled. In addition, branch road 94
Assuming that the discharge hydraulic pressure Pq generated in the downstream portion 94c of the hydraulic rotary machine increases, this discharge hydraulic pressure Pq and the seating area of the inner hole 99A of the valve body portion 181a of the valve body 181 relative to the inflow hole are When the force given by the product of is larger than the above-mentioned set load of the coil spring 182, the valve body 181 dissociates from the inflow hole of the inner hole 99A and connects the downstream portion 94c of the branch passage 94 to the pipe P9.

In such a case, under the action of the relief valve 180 as a relief valve, the discharge pressure of the hydraulic rotating machine becomes
It is maintained at a value defined by the set load of the coil spring 182. Furthermore, by controlling the inflow current J as described above, it is possible to arbitrarily obtain the discharge pressure of the hydraulic rotary machine via the coil spring 182, and when the relief valve 180 is opened, the pump capacity becomes small. , it becomes possible to downsize the relief valve 180 and control it only with the pilot section of the relief valve. In addition, linear actuator 1
By controlling the inflow current I to the relief valve 10B and the inflow current J to the relief valve 180 as described above, it is possible to obtain an arbitrary discharge pressure independently of the discharge pressure oil amount as shown in FIG. 17(A). As a result, power loss can be significantly reduced. Also,
Like the pressure compensator P/C in Fig. 18(A) which has the characteristics shown in Fig. 18(B), it is not possible to obtain independent flow characteristics with respect to changes in rotation speed,
It is also possible to reliably eliminate the problem that the operating set pressure Po cannot be controlled to an arbitrary value.

FIG. 17(B) shows a modification of the third embodiment. In this modification, the relief valve 180 in the third embodiment is replaced with a pilot of an ordinary proportional electromagnetic relief valve. Its structural feature lies in the adoption of a relief valve 180A having a structure corresponding to that of the first embodiment. In this relief valve 180A, the relief valve 18
A poppet-type valve body 189 is used in place of the valve body 181, coil spring 182, and locking member 187 of 0, and the valve body 189 is coaxially fitted to the tip of the small diameter portion 186b of the rod 186. and the conical valve body portion 18
9a, it can be seated at the inner end of the inflow hole of the inner hole 99A. The rest of the structure is the same as that of the third embodiment.

[0071] In this modified example configured in this way,
As described in the third embodiment, when a suction force is generated in the solenoid of the linear actuator 180a in response to the inflow current J flowing into the solenoid, the rod 186 absorbs the suction force and the inner hole 99A of the valve body 189. The valve body 189A is moved away from the inner end of the inflow hole of the inner hole 99A by being displaced in accordance with the difference between the seating area for the inner end of the inflow hole multiplied by the hydraulic pressure applied. Thereby, the same effects as in the third embodiment can be achieved.

[Brief explanation of the drawing]

FIG. 1 is a cross-sectional view of a hydraulic pressure control mechanism in a schematic overall cross-sectional configuration showing a first embodiment of the present invention.

FIG. 2 is a cross-sectional view of a variable displacement hydraulic rotary machine of the overall cross-sectional configuration showing the first embodiment.

[Fig. 3] A graph showing the relationship between the discharge pressure oil amount Q and capacity α of the hydraulic rotating machine and the rotation speed N, and the discharge pressure P of the hydraulic rotating machine.
FIG. 4 is a diagram showing hysteresis characteristics in the relationship between the accumulator pressure Pacc and the capacity α.

FIG. 4 is a cross-sectional view of a hydraulic pressure control mechanism in a schematic cross-sectional overall configuration showing a second embodiment of the present invention.

FIG. 5 is a cross-sectional view of a variable displacement hydraulic rotary machine of the general cross-sectional configuration showing the second embodiment.

FIG. 6 is a sectional view of a main part showing a first modification of the first or second embodiment.

FIG. 7 is a sectional view of a main part showing a partial modification of the first modification.

FIG. 8 is a diagram showing the cross-sectional shape of the inner end opening of the communication path as seen from the direction of arrow A in FIG. 7, a diagram showing each modification of the same cross-sectional shape, and a diagram showing the action on both ends of the spool of the directional control valve. 3 is a graph showing temporal changes in each hydraulic pressure after the start of unloading.

FIG. 9 is a sectional view of a main part showing a second modification of the first or second embodiment, and the relationship between the discharge pressure oil amount Q and the rotation speed N in the second modification, using the current I as a parameter. This is a graph showing.

FIG. 10 is a sectional view of a main part showing a third modification of the second embodiment.

FIG. 11 is a sectional view of a main part showing another modification of the unload valve of the first and second embodiments.

FIG. 12 is a diagram showing the overall configuration of a hydraulic rotary machine and a hydraulic pressure control mechanism according to a third embodiment of the present invention using hydraulic symbols.

13 is a schematic cross-sectional view of the proportional electromagnetic relief valve of FIG. 12. FIG.

14 is an electrical control circuit diagram for driving and controlling each linear actuator in FIG. 12. FIG.

15 is a flowchart showing the operation of the microcomputer in FIG. 14. FIG.

FIG. 16 is a graph showing the relationship between each inflow current Ib, Jb and vehicle speed V using steering angle θ as a parameter, and each inflow current I
It is a graph showing the relationship between a, Ja and maximum acceleration Gmax.

FIG. 17 is a graph showing the relationship between the discharge pressure oil amount Q and the discharge pressure Pq using each inflow current I, J as a parameter, and the third
FIG. 7 is a schematic cross-sectional view showing a modification of the embodiment.

FIG. 18 is a configuration diagram of a pressure compensator and a graph showing its characteristics.

[Explanation of symbols]

P, Pa...Variable displacement hydraulic rotating machine, 80...Reservoir tank, 110...Main throttle, 110A, 110B, 180a...
Linear actuator, 130...Auxiliary aperture, 140, 1
40A... Directional switching valve, 150, 150A... Unload valve, 180... Relief valve.

Claims (1)

[Claims] 1. Variable capacity hydraulic rotation means that is rotated by a prime mover mounted on a vehicle, sucks in hydraulic fluid in a reservoir tank according to its capacity, and discharges it to a load as pressure fluid; a check valve means that only allows discharge of pressurized fluid to the load; a main throttling means that throttles the pressurized fluid from the hydraulic rotation means to the check valve means at a predetermined low throttling degree; and this main throttling means. When the hydraulic pressure on the downstream side of the valve increases appropriately, the pressure fluid from the main throttling means is released to the reservoir tank in order to unload the hydraulic rotation means, and this released state is transferred to the downstream side of the check valve means. an unloading valve means that maintains the pressure fluid according to the side liquid pressure; an auxiliary throttling means that throttles the pressure fluid from the main throttling means to the unloading valve means at a predetermined high throttling degree; and an upstream hydraulic pressure of the main throttling means. and the downstream hydraulic pressure of the auxiliary throttling means is small, the pressure fluid downstream of the check valve means is controlled to be applied to the hydraulic rotation means to increase its capacity, and A hydraulic pressure control device for a vehicle, comprising: capacity control means for applying pressure fluid on the downstream side of the check valve means to the hydraulic rotation means to reduce its capacity when the pressure is high.