JP2008025431A - Fluid pressure actuator - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide an actuator excellent in function mounting properties by making the actuator compact in size without providing a separate load reducing means other than the actuator. <P>SOLUTION: A fluid pressure actuator generating driving force by the movement of a displacement member 54 by fluid includes a first pressure chamber 61 and a second pressure chamber 62 defined by the displacement member, and is provided with a first open close valve 67 provided in a first passage 63 establishing communication between the first pressure chamber and a pressure source 15 and a second open close valve 68 provided in a second passage 66 establishing communication between the second pressure chamber and the pressure source. The first open close valve is closed when pressure in the first pressure chamber is higher than pressure in the second pressure chamber and the second open close valve is closed when pressure in the second pressure chamber is higher than pressure in the first pressure chamber. <P>COPYRIGHT: (C)2008,JPO&INPIT

Description

  The present invention relates to an actuator, and more particularly to a fluid pressure actuator to which a variable torque is applied as an external force.

2. Description of the Related Art Conventionally, a mechanism for reducing the load load of an actuator to which fluctuating torque is applied as an external force is known. For example, in the invention disclosed in the patent document, in the variable compression ratio mechanism, a hydraulic retainer is provided as a load reducing means for reducing the fluctuation torque applied to the electric actuator, and the excessive torque applied to the electric actuator exceeds the allowable torque of the electric actuator. In such a case, the hydraulic retainer receives excessive torque to reduce the size of the electric actuator.
JP 2003-322036 A

  However, the combined use of an electric actuator and a hydraulic retainer increases the size, so when mounted on an internal combustion engine or the like, it is difficult to say that the engine mountability is excellent.

  Therefore, an object of the present invention is to provide an actuator excellent in engine mountability by reducing the size of the actuator without providing a separate load reducing means in addition to the actuator.

  The present invention solves the above problems by the following means. In addition, in order to make an understanding easy, although the code | symbol corresponding to embodiment of this invention is attached | subjected, it is not limited to this.

  A pressure chamber (60), a pressure source (15) for supplying a fluid to the pressure chamber, and the pressure chamber is divided into a first pressure chamber (61) and a second pressure chamber (62); The displacement member (54) that moves in the pressure chamber according to the pressure in the two pressure chambers, the first passage (63) that communicates the pressure source and the first pressure chamber, and the pressure source and the second pressure chamber are communicated. A second passage (66), a switching valve (73) provided upstream of the first passage and the second passage and downstream of the pressure source, and controlling supply of fluid to the first pressure chamber and the second pressure chamber; A first on-off valve (67) provided in the first passage downstream of the valve for opening and closing the first passage; a second on-off valve (68) provided in the second passage downstream of the switching valve for opening and closing the second passage; The first on-off valve is closed when the pressure in the first pressure chamber is higher than the pressure in the second pressure chamber, and the pressure in the second pressure chamber is higher than the pressure in the first pressure chamber. Second on-off valve is closed when the fluid pressure actuator displacement member generates a driving force by being moved by the fluid (8).

  According to the present invention, since the on-off valve opens and closes according to the pressure in the pressure chamber, when a large or small varying torque is applied to the fluid pressure actuator, a load opposite to the direction to be driven is applied to the fluid pressure actuator. In this case, the fluid pressure actuator functions as a cage that receives the load, and when a load is applied in the same direction as the direction of driving, the fluid actuator functions as an actuator that performs driving. It is not necessary to provide a load reducing means such as a fluid pressure actuator that is small in size and excellent in engine mountability.

  Furthermore, the load applied to the fluid pressure actuator is not received by the pressure from the pressure source, but is closed by the fluid confined in the pressure chamber by closing the on-off valve provided in the passage supplying the fluid from the pressure source to the pressure chamber. Since the load is received, there is no need to have a large pressure source to counter the maximum load. Therefore, the fluid leakage between the pressure chambers can be reduced, and the pressure source drive energy can be reduced.

  First Embodiment Hereinafter, a first embodiment of the present invention will be described with reference to the drawings.

  FIG. 1 is a configuration diagram of a variable compression ratio mechanism and a variable valve mechanism of an engine according to the first embodiment of the present invention.

  Reference numeral 1 denotes a variable compression ratio mechanism that makes the compression ratio of an in-line four-cylinder engine variable. 2 is a piston, 3 is a crankshaft.

  The crankshaft 3 includes a plurality of journals 3a and a crankpin 3b. The journal 3 a is rotatably supported by the main bearing of the cylinder block 13. The crank pin 3b is eccentric from the journal 3a by a predetermined amount, and a lower link 5 serving as a second link is rotatably connected thereto.

  The lower link 5 is configured to be split into two left and right members, and a crank pin 3b is fitted in a substantially central connecting hole.

  The upper link 6 serving as the first link has a lower end side rotatably connected to one end of the lower link 5 by a first connecting pin 11, and an upstream side rotatably connected to the piston 2 by a piston pin 4.

  The piston 2 receives combustion pressure and reciprocates in the cylinder 14 of the cylinder block 13.

  The control link 7 serving as the third link is pivotally connected to the other end of the lower link 5 at the upper end side by the second connecting pin 12, and the cylinder block 13 whose lower end side is part of the engine body via the control shaft 9. It is connected to the lower part of the pivotably.

  FIG. 2 is a side view showing a configuration near the control shaft 9. The cross section is hatched to facilitate understanding of the shaft penetration state.

  As shown in FIG. 2, an eccentric sleeve bearing 50 is provided at the lower end side of the control link 7, and the eccentric shaft portion 10 of the control shaft 9 is rotatably inserted into the eccentric sleeve bearing 50.

  As shown in FIG. 1, a lever 51 is provided on the control shaft 9, and a sliding pin 52 is slidably fitted in a slit 51 a formed in the lever 51. One end of a rod 53 is connected to the slide pin 52, and the other end of the rod 53 is connected to an actuator piston 54 of the hydraulic actuator 8.

  When the rod 53 is driven by the hydraulic actuator 8, the control shaft 9 rotates, and the center position of the eccentric shaft portion 10, particularly the relative position with respect to the engine body changes. As a result, the motion constraint condition of the lower link 5 by the control link 7 changes, the piston stroke characteristics including the piston top dead center position with respect to the crank angle change, and the engine compression ratio is changed.

  In the present embodiment, as shown in FIG. 3 (a), the control shaft 9 rotates in the direction A1 (clockwise), and the tip of the rod 53 provided with the sliding pin 52 moves away from the hydraulic actuator 8. (A2 direction) is set to be the low compression ratio direction.

  Therefore, at the time of low compression ratio control, as shown in FIG. 3A, the control shaft 9 is rotated clockwise together with the lever 51 by the rod 53, whereby the control link 7 is raised upward by the eccentric shaft portion 10. Then, the lower link 5 moves counterclockwise and the first connecting pin 11 is lowered, so that the position of the top dead center of the piston 2 is lowered.

  On the other hand, at the time of high compression ratio control, as shown in FIG. 3B, the control shaft 9 is rotated counterclockwise by the rod 53 together with the lever 51, whereby the control link 7 is moved downward by the eccentric shaft portion 10. The lower link 5 is moved clockwise and the first connecting pin 11 is raised, so that the position of the top dead center of the piston 2 is raised.

  3A and 3B representatively show the low compression ratio state and the high compression ratio state, but it is possible to continuously change the compression ratio between them. is there.

  As shown in FIG. 1, oil is supplied from the oil pump 15 to the hydraulic actuator 8. The supply of oil to the hydraulic actuator 8 is adjusted by a solenoid valve 17. The oil pump 15 also supplies hydraulic pressure to a later-described variable valve mechanism 20 via the control valve 18.

  A control unit (ECU) 16 is provided to control the variable compression ratio mechanism 1 and the variable valve mechanism 20, and the ECU 16 adjusts values such as engine speed, engine load, suction negative pressure, and hydraulic pressure obtained from various sensors. Based on this, the oil pump 15, the solenoid valve 17, and the control valve 18 are controlled.

  The variable valve mechanism 20 is a variable valve mechanism that can continuously change the intake valve lift characteristics, and continuously changes the valve lift amount and the operating angle of the intake valve by changing the rotation angle of the control shaft 23. And a second variable valve mechanism 22 capable of continuously changing the center phase of the operating angle of the intake valve to the advance side and the retard side.

  In response to a command from the ECU 16, hydraulic pressure is supplied to the first actuator 24 that drives the control shaft 23 from the oil pump 15 and the second actuator (not shown) built in the second variable valve mechanism 22 via the control valve 18. The intake valve lift characteristic is controlled.

  The configuration of the first variable valve mechanism 21 will be described with reference to FIGS.

  In the first variable valve mechanism 21, a control shaft 23 and a drive shaft 25 are rotatably supported by a cylinder head 27 via bearing brackets 26 (see FIGS. 4 and 5).

  A swing cam 28 for driving the intake valve is externally fitted to the drive shaft 25 so as to be rotatable for each cylinder. A shim or the like is provided between each swing cam 28 and the valve stem 29 of the intake valve. A solid type valve lifter 30 for adjusting the valve clearance is interposed.

  As shown in FIGS. 1 and 3, the swing cam 28 includes a pair of cam main bodies 28 a that are in contact with a pair of valve lifters 30 of each cylinder by a cylindrical journal portion 28 b that is rotatably supported by a bearing bracket 26. The structure is integrally connected.

  A cylindrical eccentric shaft portion 31 is fixed or integrally formed on the drive shaft 25, and the axis X of each eccentric shaft portion 31 is eccentric with respect to the rotation center Y of the drive shaft 25.

A ring-shaped first link 32 is rotatably fitted on the outer periphery of the eccentric shaft portion 31. That is, the first link 32 is supported so as to be rotatable around the axis Y of the eccentric shaft portion 31.
A cylindrical control eccentric shaft portion 33 is fixed or integrally formed on the control shaft 23 for each cylinder, and the axis P1 of the control eccentric shaft portion 33 is eccentric with respect to the rotation center P2 of the control shaft 23. .

  A central portion of the rocker arm 34 is rotatably fitted on the outer periphery of the control eccentric shaft portion 33. That is, the rocker arm 34 is supported so as to be rotatable around the axis P <b> 1 of the control eccentric shaft portion 33.

  One end of the rocker arm 34 is rotatably connected to the tip of the first link 32 via the first connecting pin 35, and the other end of the rocker arm 34 is connected to one end of the rod-like second link 36 with the second connecting pin 37. It is connected via a rotation. The other end of the second link 36 is connected to the tip of the swing cam 28 via a third connecting pin 38.

  When the drive shaft 25 rotates in conjunction with the rotation of the crankshaft, the axis X of the eccentric shaft portion 31 serving as the rotation center of the first link 32 turns with respect to the rotation center Y of the drive shaft 25, and this first link. 32, the rocking cam 28 rocks within a predetermined angle range via the rocker arm 34 and the second link 36.

  The cam surface 39 of the swinging swing cam 28 comes into contact with and presses the valve lifter 30, whereby the valve stem 29 of the intake valve is pushed down via the valve lifter 30, and the intake valve is opened and closed. .

  Further, when the control shaft 23 is controlled to rotate according to the engine operating state via the first actuator 24, the axis P1 of the control eccentric shaft portion 33 serving as the rocking center of the rocker arm 34 becomes the rotation center P2 of the control shaft 23. And the postures of the rocker arm 34 and the links 32 and 36 change. As a result, the swing characteristic of the swing cam 28 changes, and the intake valve lift characteristic changes.

  Specifically, the intake valve operating angle and the valve lift amount continuously change while the phase of the operating angle of the intake valve with respect to the rotational phase of the crankshaft 3 remains substantially constant.

  The second variable valve mechanism 22 operates the intake valve by relatively changing the phase of the drive shaft 25 with respect to the pulley or sprocket 40 rotating in conjunction with the crankshaft 3 via the second actuator (not shown). The phase of the angle can be continuously changed, and a helical spline type, a vane type, and the like are known, and detailed description thereof is omitted.

  By providing the first variable valve mechanism 21 and the second variable valve mechanism 22 as described above, the valve lift amount, the operating angle, and the operating angle phase can be continuously changed, and the optimal valve lift for the operating state. Characteristics can be realized.

  Then, the structure and operation | movement are demonstrated based on FIG. 6 which is a block diagram of the hydraulic actuator 8 in 1st embodiment of this invention.

  The hydraulic actuator 8 has a cylindrical cylinder 60, and the inside of the cylinder 60 is partitioned into a first hydraulic chamber 61 and a second hydraulic chamber 62 by an actuator piston 54. The actuator piston 54 can reciprocate in the axial direction of the cylinder 60, and the actuator piston 54 is displaced according to changes in pressure in the first hydraulic chamber and the second hydraulic chamber 62.

  A rod 53 is connected to the actuator piston 54, and a sliding pin 52 is provided at the other end of the rod 53. The sliding pin 52 is slidably connected to a slit 51 a of a lever 51 provided on the control shaft 9. Therefore, when the displacement of the actuator piston 54 is changed, the control shaft 9 rotates and the engine compression ratio can be changed.

  In order to control the hydraulic pressure supplied to the hydraulic actuator 8, a hydraulic pressure switching valve 65 driven by a solenoid 17 is provided. A first hydraulic chamber 61 of the hydraulic actuator 8 is connected to the hydraulic switching valve 65 by a first passage 63, and a second hydraulic chamber 62 is connected by a second passage 66.

  A first on-off valve 67 is provided in the middle of the first passage 63, and a second on-off valve 68 is provided in the second passage 66. The first on-off valve 67 and the second on-off valve 68 are configured as pilot-type operation valves.

  One end of the first on-off valve 67 is urged by a first spring 69 in the direction in which the first on-off valve 67 is closed, and the other end is communicated with the second hydraulic chamber 62 by a first pilot passage 71. That is, the first on-off valve 67 operates using the hydraulic pressure of the second hydraulic chamber 62 guided to the first pilot passage 71 as the pilot pressure.

  When the hydraulic pressure in the first pilot passage 71 (or the second hydraulic chamber 62) is equal to or greater than a predetermined value, the hydraulic pressure becomes larger than the urging force of the first spring 69, the first on-off valve 67 is opened, and the first When the hydraulic pressure in the pilot passage 71 (or the second hydraulic chamber 62) is less than a predetermined value, the urging force of the first spring 69 is greater than the hydraulic pressure, so the first on-off valve 67 is closed and the first hydraulic chamber is closed. It functions as a check valve that prevents the flow of oil from the 61 side to the hydraulic pressure switching valve 65 side.

  Similarly, one end of the second on-off valve 68 is urged by the second spring 70 in the closing direction of the second on-off valve 68, and the other end communicates with the first hydraulic chamber 61 by the second pilot passage 72. . That is, the second on-off valve 68 operates using the hydraulic pressure of the first hydraulic chamber 61 guided to the second pilot passage 72 as the pilot pressure.

  When the hydraulic pressure in the second pilot passage 72 (or the first hydraulic chamber 61) is equal to or greater than a predetermined value, the hydraulic pressure becomes larger than the urging force of the second spring 70, the second on-off valve 68 is opened, and the second When the hydraulic pressure in the pilot passage 72 (or the first hydraulic chamber 61) is less than a predetermined value, the urging force of the second spring 70 is larger than the hydraulic pressure, so the second on-off valve 68 is closed and the second hydraulic chamber is closed. It functions as a check valve that prevents the flow of oil from the 62 side to the hydraulic pressure switching valve 65 side.

  The oil pressure switching valve 65 is provided with a slidable spool valve 73 driven by a solenoid 17. The solenoid 17 is energized and controlled by the ECU 16. One end of the spool valve 73 is pushed by the plunger 17 a of the solenoid 17, and the other end is biased by the spring 74. That is, the position of the spool valve 73 is determined by the position of the plunger 17 a that is driven by energization of the solenoid 17.

  The hydraulic switching valve 65 includes a port A connected to the first passage 63 and a port B connected to the second passage 66. Furthermore, discharge ports C and D connected to an oil pan (not shown) are provided.

  The spool valve 73 is controlled to the three position by the solenoid 17. The third position is a neutral position where the spool valve 73 closes both the ports A and B, an A position where the spool valve 73 is moved toward the port A side, and a B position where the spool valve 73 is moved toward the port B side. In the A position, port A communicates with a pump port 64 connected to the oil pump 15, and port B communicates with a discharge port D at the same time. In the B position, port B communicates with pump port 64 and port A communicates with discharge port C. In other words, the spool valve 73 functions as a three-way switching valve, and distributes whether the oil supplied from the oil pump 15 is supplied to the first hydraulic chamber 61 or the second hydraulic chamber 62, or neither. ing. Hereinafter, the flow of oil at each position will be described.

  When the spool valve 73 is in the neutral position, the spool valve 73 blocks both the port A and the port B, so that the oil from the oil pump 15 is blocked by the spool valve 73, and no oil is supplied to the cylinder 60. . Further, oil is not discharged from the cylinder 60.

  When the spool valve 73 is in the A position, the spool valve 73 is moved toward the port A, and at this time, the pump port 64 and the port A are communicated with each other. Accordingly, the oil supplied from the oil pump 15 passes through the first passage 63 through the port A from the pump port 64 and reaches the first on-off valve 67. Here, when the first on-off valve 67 is open, the oil is supplied to the first hydraulic chamber 61, and when the first on-off valve 67 is closed, the oil is supplied / discharged in the first hydraulic chamber 61. Is not done.

  Further, when the spool valve 73 is in the A position, the discharge port D and the port B are communicated. As described above, since the oil is supplied from the oil pump 15 to the first hydraulic chamber 61 when the first on-off valve 67 is open, the actuator piston 54 is moved to the second hydraulic chamber by the hydraulic pressure of the first hydraulic chamber 61. It is pushed to 62 side. Accordingly, the oil in the second hydraulic chamber 62 is pushed out to the second passage 66 and reaches the second on-off valve 68. Here, when the second on-off valve 68 is open, the oil pushed out from the second hydraulic chamber 62 passes through the second passage 66, passes through the port B, and is discharged from the discharge port D into an oil pan (not shown). Is discharged. When the second on-off valve 68 is closed, the oil in the second hydraulic chamber 62 is not supplied / discharged.

  Next, when the spool valve 73 is in the B position, the spool valve 73 is moved to the port B side, and at this time, the pump port 64 and the port B are communicated. Therefore, the oil supplied from the oil pump 15 passes through the second passage 66 through the port B from the pump port 64 and reaches the second on-off valve 68. Here, when the second on-off valve 68 is open, the oil is supplied to the second hydraulic chamber 62, and when the second on-off valve 68 is closed, the oil is supplied / discharged in the second hydraulic chamber 62. Is not done.

  Further, when the spool valve 73 is in the B position, the discharge port C and the port A are communicated. As described above, when the second on-off valve 68 is open, oil is supplied from the oil pump 15 to the second hydraulic chamber 62, so that the actuator piston 54 is moved to the first hydraulic chamber by the hydraulic pressure of the second hydraulic chamber 62. It is pushed to 61 side. Accordingly, the oil in the first hydraulic chamber 61 is pushed out to the first passage 63 and reaches the first on-off valve 67. Here, when the first on-off valve 67 is open, the oil pushed out from the first hydraulic chamber 61 passes through the first passage 63, passes through the port A, and is discharged from the discharge port C to an oil pan (not shown). Is discharged. When the first on-off valve 67 is closed, the oil in the first hydraulic chamber 61 is not supplied / discharged.

  Next, control of the control shaft 9 by the hydraulic actuator 8 will be described with reference to FIG.

  The control shaft 9 is subjected to a load in which the in-cylinder pressure and the link inertia force are combined, and a torque combined with the amount of eccentricity of the control shaft (hereinafter referred to as “control shaft torque”) is applied. As shown in FIG. 7, the control shaft torque is a fluctuating torque in which large and small torques appear periodically, and the torque is large in the direction in which the in-cylinder pressure acts on the piston, that is, in the low compression ratio direction of the compression ratio (ε). Become. Since this example is a four-cylinder engine, the in-cylinder pressure is applied twice during one rotation of the crankshaft 3 (during a crank angle of 360 degrees), and the maximum torque is generated twice in the direction of the low compression ratio.

  Accordingly, when the control shaft 9 is rotated in the high compression ratio direction, the hydraulic actuator 8 needs to overcome the maximum control shaft torque in the low compression ratio direction to drive the lever 51, and an excessive pressure receiving surface ( The surface perpendicular to the sliding direction of the actuator piston 54 and the surface perpendicular to the sliding direction of the actuator piston 54 of the cylinder 60) and hydraulic pressure are required, and the hydraulic actuator 8 and the hydraulic source (oil pump 15) are increased in size.

  The gist of the present invention is that the control shaft 9 can be driven without having an excessive pressure receiving surface or a hydraulic pressure source, or without providing a mechanism for reducing the control shaft torque separately from the actuator. 8 is to prevent an increase in size.

  Hereinafter, control of the control shaft 9 will be described with reference to FIG.

  The lower part of FIG. 7 shows the rotation angle of the control shaft with respect to the crank angle when the control shaft 9 is rotated in the high compression ratio direction. As shown in the lower part of FIG. 7, when the control shaft is rotated in the high compression ratio direction, the control shaft is rotated only when the control shaft torque in the high compression ratio direction is generated, and the control shaft torque in the low compression ratio direction is When this occurs, the control shaft 9 is not rotated but is held at that position.

  That is, when the hydraulic actuator 8 rotates the control shaft 9 in the high compression ratio direction, if the control shaft torque in the same direction (high compression ratio direction) is generated, the hydraulic actuator 8 functions as an actuator. On the other hand, when a control shaft torque having a large torque in the direction in which the control shaft 9 rotates in the reverse direction (low compression ratio direction) is generated, the hydraulic actuator 8 functions as a cage.

  On the other hand, when the control shaft torque in the low compression ratio direction is generated when the control shaft 9 is rotated in the low compression ratio direction, the direction of the load acting on the hydraulic actuator 8 and the direction in which the hydraulic actuator 8 is desired to operate. Therefore, it is possible to assist the sliding of the actuator piston 54 using the control shaft torque with a large torque in the low compression ratio direction, and to rapidly rotate the control shaft 9. On the other hand, when the control shaft torque in the high compression ratio direction is generated, the control shaft 9 is not rotated but held at the position at that time.

  That is, when the hydraulic actuator 8 rotates the control shaft 9 in the low compression ratio direction, if the control shaft torque in the same direction (low compression ratio direction) is generated, the hydraulic actuator 8 functions as an actuator. On the other hand, when the control shaft torque is generated in the direction in which the control shaft 9 is rotated in the reverse direction (high compression ratio direction), the hydraulic actuator 8 functions as a cage.

  As described above, the absolute value of the torque is greater when the control shaft torque acts in the low compression ratio direction (negative torque) than when it acts in the high compression ratio direction (positive torque). Therefore, when performing the control as described above, the hydraulic actuator 8 only needs to have an allowable load that can hold the control shaft 9 with respect to the maximum control shaft torque in the low compression ratio direction.

  Next, a case where the hydraulic actuator 8 is operated to the high compression ratio side of the engine compression ratio will be described with reference to FIG.

  Changing the compression ratio of the hydraulic actuator 8 to the high compression ratio side means that the actuator piston 54 is slid to the second hydraulic chamber 62 side as shown in FIG. Therefore, the ECU 16 controls the solenoid 17 so that the spool valve 73 of the hydraulic pressure switching valve 65 is moved to the A position (port A side) in order to supply the pressure oil from the oil pump 15 to the first hydraulic chamber 61.

  As described above, the torque changes in the control shaft 9 in different directions at a rate of twice per engine rotation, and when the control shaft 9 is rotated in the high compression ratio direction, the same direction (high When the control shaft torque is generated in the compression ratio direction), the hydraulic actuator 8 functions as an actuator and rotates the control shaft 9. On the other hand, when the control shaft torque is generated in the direction opposite to the direction of rotation (low compression ratio direction), the hydraulic actuator 8 functions as a cage and holds the control shaft 9 at that position.

  FIG. 8 (a) shows a case where the control shaft torque is applied in the same direction (high compression ratio direction) as the direction of rotation when the control shaft 9 is rotated in the high compression ratio direction in order to increase the engine compression ratio. 2 is a diagram of a hydraulic actuator 8. FIG.

  When the control shaft torque in the high compression ratio direction is applied, a load is applied to the rod 53 and the actuator piston 54 in the direction indicated by the arrow F1, that is, the direction in which the tip of the rod 53 provided with the sliding pin 52 approaches the hydraulic actuator 8. Take it.

  At this time, the second hydraulic chamber 62 has a higher pressure than the first hydraulic chamber 61, and the hydraulic pressure applied to the first on-off valve 67 via the first pilot passage 71 communicating with the second hydraulic chamber 62 is equal to or greater than a predetermined value. Thus, the urging force of the first spring 69 is overcome and the first on-off valve 67 is opened. When the first on-off valve 67 is opened, oil is supplied from the oil pump 15 to the first hydraulic chamber 61, and the actuator piston 54 is pushed toward the second hydraulic chamber.

  Therefore, the pressure in the first hydraulic chamber 61 rises, the hydraulic pressure applied to the second on-off valve 68 via the second pilot passage 72 overcomes the urging force of the second spring 70, and the second on-off valve 68 is opened. Accordingly, the actuator piston 54 moves while discharging the oil in the second hydraulic chamber, and the rod 53 also moves together with it, and the lever 51 and the control shaft 9 are rotated in the high compression ratio direction.

  FIG. 8 (b) shows a case where the control shaft torque in the direction opposite to the direction of rotation (low compression ratio direction) is applied when the control shaft 9 is rotated in the high compression ratio direction in order to increase the engine compression ratio. 2 is a diagram of a hydraulic actuator 8. FIG. When a large control shaft torque in the low compression ratio direction is applied, the rod 53 and the actuator piston 54 are greatly increased in the direction indicated by the arrow F2, that is, in the direction in which the tip of the rod 53 provided with the sliding pin 52 moves away from the hydraulic actuator 8. A load is applied. Accordingly, the actuator piston 54 tends to be moved toward the first hydraulic chamber 61 (in the low compression ratio direction), but the pressure in the second hydraulic chamber 62 becomes less than a predetermined value, and the biasing force of the first spring 69 is the first. Since the hydraulic pressure applied to the first on-off valve 67 is overcome via the pilot passage 71, the first on-off valve 67 is closed.

  For this reason, since the first hydraulic chamber 61 is sealed against the force pulling the actuator piston 54 toward the first hydraulic chamber 61 via the rod 53, a high compression reaction force is generated, and the actuator piston 54 is Can be held in that position.

  The operations of (a) and (b) of FIG. 8 described above are repeated while the hydraulic switching valve 65 is switched to the A position. During this time, the actuator piston 54 strokes to the high compression ratio side by a predetermined amount. When the predetermined high compression ratio is reached, the hydraulic switching valve 65 is returned to the neutral position by the ECU 16. The compression ratio is maintained in that state.

  Next, a case where the hydraulic actuator 8 is operated to the low compression ratio side will be described with reference to FIG.

  Changing the hydraulic actuator 8 to the low compression ratio side means that the actuator piston 54 is moved to the first hydraulic chamber 61 side. Accordingly, the ECU 16 controls the solenoid 17 so that the spool valve 73 can be moved to the B position (port B side) in order to supply the oil from the oil pump 15 to the second hydraulic chamber 62.

  As described above, when the control shaft 9 is rotated in the low compression ratio direction, if the control shaft torque is generated in the same direction (low compression ratio direction) as the direction in which the control shaft 9 is to be rotated, the hydraulic actuator 8 is used as an actuator. Functions and rotates the control shaft 9. Further, since the torque in the low compression ratio direction is large, the sliding of the actuator piston 54 can be assisted by using this torque, and the control shaft 9 can be quickly rotated. On the other hand, when the control shaft torque is generated in the direction opposite to the direction of rotation (high compression ratio direction), the hydraulic actuator 8 functions as a cage, and the control shaft 9 is not rotated and held at that position. To do.

  FIG. 9A is a diagram of the hydraulic actuator 8 when a large control shaft torque is applied in the low compression ratio direction when the control shaft 9 is rotated in the low compression ratio direction in order to lower the engine compression ratio. When the control shaft torque is applied in the low compression ratio direction, a load is applied to the rod 53 and the actuator piston 54 in the direction indicated by the arrow F 3, that is, the direction in which the tip of the rod 53 provided with the sliding pin 52 moves away from the hydraulic actuator 8. Take it.

  At this time, since the direction in which the actuator piston 54 is desired to be moved coincides with the direction in which the control shaft torque is applied, the hydraulic actuator 8 can easily reduce the engine compression ratio. That is, since the hydraulic pressure in the first hydraulic chamber 61 becomes equal to or greater than a predetermined value, the hydraulic pressure applied to the second on-off valve 68 via the second pilot passage 72 overcomes the urging force of the second spring 70, and the second on-off valve 68 is be opened. When the second on-off valve 68 is opened, oil is supplied from the oil pump 15 to the second hydraulic chamber 62, and the actuator piston 54 is pushed toward the first hydraulic chamber. As the pressure in the second hydraulic chamber 62 rises, the first on-off valve 67 is opened, so that the actuator piston 54 moves while discharging the oil in the first hydraulic chamber 61, and the rod 53 also moves together with it. 51 and the control shaft 9 are rotated in the high compression ratio direction.

  FIG. 9B is a diagram of the hydraulic actuator 8 when a small control shaft torque in the high compression ratio direction is applied when the control shaft 9 is rotated in the low compression ratio direction in order to lower the engine compression ratio. When the control shaft torque in the high compression ratio direction is applied, a load is applied to the rod 53 and the actuator piston 54 in the direction indicated by the arrow F4, that is, the direction in which the tip of the rod 53 provided with the sliding pin 52 approaches the hydraulic actuator 8. Take it.

  At this time, the actuator piston 54 tends to be pushed back toward the second hydraulic chamber 62 (in the high compression ratio direction). However, since the hydraulic pressure in the first hydraulic chamber 61 becomes less than a predetermined value, the urging force of the second spring is the first. The hydraulic pressure applied to the second on-off valve 68 is overcome via the two pilot passages 72, and the second on-off valve 68 remains closed.

  For this reason, the second hydraulic chamber 62 is sealed, and the actuator piston 54 is held at the position facing the load F4 by the compression reaction force. Thereafter, while the hydraulic pressure switching valve 65 is switched to the B position, the operations (A) and (B) in FIG. 9 are repeated, and the actuator piston 54 strokes to a position where a predetermined compression ratio is obtained.

  When the compression ratio is maintained without being changed, the ECU 16 controls the solenoid 17 so that the spool valve 73 is in the neutral position in order to hold the control shaft 9 in that position. At this time, the hydraulic actuator 8 functions as a cage.

  According to this embodiment, the first opening / closing valve is provided in the first passage, the second opening / closing valve is provided in the second passage, and the first opening / closing valve is provided when the pressure in the first hydraulic chamber is higher than the pressure in the second hydraulic chamber. Since the second on-off valve is closed when the pressure in the first hydraulic chamber is higher than the pressure in the second hydraulic chamber, the on-off valve is closed even when the hydraulic actuator 8 receives a load greater than the hydraulic pressure of the hydraulic source. By closing, the load can be received by the oil trapped in the hydraulic chamber. Further, oil passes through the passage having the open on-off valve, and oil is supplied to the pressure chamber communicated with the passage, so that the actuator piston 54 is slid to generate a driving force. Therefore, the hydraulic actuator 8 has functions of both an actuator and a cage, and it is not necessary to provide another cage for reducing the load on the actuator, and is excellent in engine mounting.

In particular, in this embodiment, since the hydraulic actuator 8 of the present invention is applied to drive the control shaft 9 of the multi-link variable compression ratio mechanism 1, the hydraulic pressure of the hydraulic source in the low compression ratio direction applied to the control shaft 9 Can be received by the hydraulic actuator 8, so that the control shaft 9 is rotated only when the direction in which the control shaft 9 is to be rotated matches the direction of the control shaft torque. When the control shaft torque in the reverse direction is applied, the control shaft can be held and reverse rotation can be prevented, so that the output required for the hydraulic actuator 8 can be kept small. Realized low pressure of (oil pump 15) It is.

  Furthermore, by realizing a reduction in the size of the hydraulic actuator 8 and a reduction in the pressure of the hydraulic source, oil leakage between the first hydraulic chamber 61 and the hydraulic chamber 62 can be reduced, and fuel consumption deteriorates due to a reduction in hydraulic source drive energy. Can be prevented.

  In addition, the first hydraulic chamber 61 and the second on-off valve 68 are communicated with each other by the first pilot passage 71, and the second hydraulic chamber 62, the first on-off valve 67, and the second pilot passage 72 are communicated with each other. Since the 67 and the second on-off valve 68 are driven by the hydraulic pressure of the hydraulic chamber, the actuator can be downsized and the hydraulic source drive energy can be reduced with a simple configuration.

  In addition, since the hydraulic pressure source can be reduced, the hydraulic pressure source for driving the hydraulic actuator 8 and the variable valve mechanism 20 can be shared, and a plurality of hydraulic pressure sources are not required, so that a simple system can be realized. .

Further, since the hydraulic actuator 8 can be driven at a low hydraulic pressure, when the engine oil pump is shared between the hydraulic actuator 8 and the variable valve mechanism 20, the engine speed is low during idling and the hydraulic pressure is low. In this case, the responsiveness of the variable movement mechanism 20 can be kept good.
(Second Embodiment)
Hereinafter, a second embodiment of the present invention will be described with reference to FIG.

  In the present embodiment, the configuration of the actuator piston 54 is different from that of the first embodiment. FIG. 10 is a configuration diagram of the hydraulic actuator 8 in the second embodiment.

  A piston spring 80 is provided in the first pressure chamber 61 so as to bias the actuator piston 54 in the high compression ratio direction.

As described above, the control shaft torque applied to the hydraulic actuator 8 is a variable torque as shown in FIG. 7, and is large in the low compression ratio direction. Therefore, according to the present embodiment, since the piston spring 80 is provided so as to bias the actuator piston 54 in the high compression ratio direction, the control shaft torque that increases in the low compression ratio direction can be generated by the biasing force of the piston spring 80. The control shaft 9 can be driven with a lower pressure source.
(Third embodiment)
Hereinafter, a third embodiment of the present invention will be described with reference to FIG.

  In the present embodiment, the first on-off valve 67 and the second on-off valve 68 are different from the above-described embodiment. FIG. 11 is a configuration diagram of the hydraulic actuator 8 in the third embodiment.

  81 is a first cam and 82 is a second cam. The first on-off valve 67 is driven by the first cam 81, and the second on-off valve 68 is driven by the second cam 82. A first spring 69 is provided at one end of the first on-off valve 67, and a second spring 70 is provided at one end of the second on-off valve 68 so as to urge the valve in the opening direction. The first cam 81 is provided with two first cam peaks 81a, and the second cam 82 is provided with two second cam peaks 82a, and the first on-off valve 67 is provided with the first cam peaks 81a. When the second cam crest 82a pushes the on-off valve 68, each valve is closed. When in the base circle of the first cam 81 and the second cam 82, that is, when the first cam crest 81a and the second cam crest 82a are at positions shifted from the first on-off valve 67 and the second on-off valve 68, The first and second on-off valves 67 and 68 are held at the open positions by the biasing force of the springs.

  The first cam 81 and the second cam 82 are moved in synchronization with the crank. As described above, the torque applied to the control shaft 9 is a fluctuation torque in which large and small positive and negative torques appear periodically as shown in FIG. 7, and in the direction in which the in-cylinder pressure acts on the piston, that is, in the low compression ratio direction. Large torque. Since the present embodiment is a four-cylinder engine, the in-cylinder pressure acts twice during one rotation of the crank (crank angle of 360 degrees), and the maximum torque is generated twice. The maximum torque in the high compression ratio direction is also generated twice during one crank rotation (crank angle of 360 degrees).

  Therefore, the first cam 81 and the second cam 82 are each provided with two cam peaks, and the control shaft torque on each side in the high compression ratio direction and the low compression ratio direction is greater than the torque generated by the hydraulic pressure generated by the oil pump 15. On the other hand, the on-off valves 67 and 68 are closed during a longer period. That is, in a crank angle period in which a torque greater than a predetermined value in the high compression ratio direction is generated, the second cam peak 82 closes the second on-off valve 68, and in a crank angle period in which a torque greater than a predetermined value in the low compression ratio direction is generated. The first cam crest 81 a closes the first on-off valve 67.

According to the present embodiment, the first on-off valve 67 and the second on-off valve 68 are mechanically driven by a cam synchronized with the crank, not hydraulically, so the length of the oil passage (pipe), the bending, Responsiveness at high rotation is improved without considering rigidity. Therefore, the responsiveness of the first on-off valve 67 and the second on-off valve 68 with respect to fluctuations in the control shaft torque is improved, and oil leakage from the pressure chamber on the relatively high pressure side can be reliably prevented. Therefore, it is possible to prevent deterioration in fuel consumption due to energy reduction.
(Fourth embodiment)
Hereinafter, a fourth embodiment of the present invention will be described with reference to FIG.

  In the present embodiment, the first on-off valve and the second on-off valve are different from the above-described embodiment. FIG. 12 is a configuration diagram of the hydraulic actuator 8 according to the fourth embodiment.

  85 is a first motor, and 86 is a second motor. The first on-off valve 67 and the second on-off valve 68 are rotary valves. The first on-off valve 67 is driven by the first motor 85 and the second on-off valve 68 is driven by the second motor 86, respectively. The first motor 85 and the second motor 86 are controlled in synchronism with crank rotation.

  In each of the embodiments, the second on-off valve 68 is closed when the control shaft torque in the high compression ratio direction is generated, and the first on-off valve 67 is closed when the control shaft torque in the low compression ratio direction is generated. In this embodiment, the on-off valve is controlled not only in the direction in which the control shaft torque is applied, but also in accordance with whether the engine compression ratio is changed to the high compression ratio direction or the low compression ratio direction. 67 and 68 are controlled.

  That is, when the control shaft torque is applied in the same low compression ratio direction as the rotation direction, such as when the engine compression ratio is changed to the low compression ratio side, the control shaft torque is applied in the direction in which the actuator piston 54 is to be moved. Since the directions match, both the first on-off valve 67 and the second on-off valve 68 are opened so that the assist by the control shaft torque is efficiently performed.

  When the control shaft 9 is rotated in the low compression ratio direction, when the control shaft torque is applied in the same low compression ratio direction as the direction in which the control shaft torque is to be rotated, both the first on-off valve 67 and the second on-off valve 68 are opened. Then, the movement of the actuator piston 54 that is about to slide toward the first hydraulic chamber 61 by the hydraulic pressure supplied from the hydraulic source to the second hydraulic chamber 62 and the control shaft torque is changed to the hydraulic pressure in the first hydraulic chamber 61. Therefore, the engine compression ratio is rapidly changed to the low compression ratio side.

  On the other hand, when the control shaft 9 is rotated in the high compression ratio direction, if the control shaft torque is applied in the same high compression ratio direction as the direction in which the control shaft torque is to be rotated, both the first on-off valve 67 and the second on-off valve 68 are turned on. As a result, the assist by the control shaft torque can be efficiently performed.

  Note that the engine compression ratio is changed not only when the control shaft torque is applied in the same low compression ratio direction as that in which the control shaft torque is to be rotated, but also when applied in the reverse rotation high compression ratio direction. The first on-off valve 67 and the second on-off valve 68 may be opened. This is because the control shaft torque in the high compression ratio direction is smaller than the control shaft torque in the low compression ratio direction, as shown in FIG.

  On the other hand, when changing the engine compression ratio in the high compression ratio direction, if both the first on-off valve 67 and the second on-off valve 68 are opened when the control shaft torque is applied in the low compression ratio direction, the low compression ratio direction Since the control shaft 9 is rotated in the low compression direction by the large control shaft torque, the first on-off valve 67 and the second on-off valve only when the control shaft torque is applied in the same high compression ratio direction as the direction in which the control shaft torque is to be rotated. 68 is limited to open.

  According to the present embodiment, when the control shaft torque is applied in the same direction as the direction in which the engine compression ratio is to be changed, both the first on-off valve 67 and the second on-off valve 68 are opened. Variable is possible. In particular, if you want to change the engine compression ratio in the low compression ratio direction, you can quickly change the compression ratio with a large control shaft torque in the low compression ratio direction, so you want to quickly reduce the compression ratio as a measure against knocking. It is effective in some cases.

  The present invention is not limited to the embodiment described above, and various modifications and changes can be made within the scope of the technical idea, and it is obvious that these are equivalent to the present invention.

  For example, in the above embodiment, oil is used as the fluid that drives the fluid pressure actuator of the present invention, but other incompressible fluids may be used. In the above embodiment, the fluid pressure actuator of the present invention is applied to drive the control shaft of the variable compression ratio mechanism. However, the present invention can also be applied to other mechanisms such as a compressor that are subject to variable torque.

It is a block diagram of the variable compression ratio mechanism and variable valve mechanism in 1st Embodiment. It is a side view of the vicinity of the control shaft of the variable compression ratio mechanism in the first embodiment. It is a state figure of the variable compression ratio mechanism in a 1st embodiment. It is a block diagram of the variable valve mechanism in 1st Embodiment. It is a side view of the variable valve mechanism in 1st Embodiment. It is a block diagram of the hydraulic actuator in 1st Embodiment. It is a graph which shows a control shaft torque and a control shaft rotation angle. It is a state figure showing the variable characteristic to the high compression ratio direction of the hydraulic actuator in a 1st embodiment. It is a state figure which shows the variable characteristic to the low compression ratio direction of the hydraulic actuator in 1st Embodiment. It is a block diagram of the hydraulic actuator in 2nd Embodiment. It is a block diagram of the hydraulic actuator in 3rd Embodiment. It is a block diagram of the hydraulic actuator in 4th Embodiment.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Variable compression ratio mechanism 2 Piston 3 Crankshaft 3a Crank journal 3b Crankpin 4 Piston pin 5 Lower link 6 Upper link 7 Control link 8 Hydraulic actuator 9 Control shaft 10 Eccentric shaft part 11 First connection pin 12 Second connection pin 13 Cylinder Block 14 Cylinder 15 Oil pump 16 ECU
17 Solenoid 18 Control valve 20 Variable valve mechanism 21 First variable valve mechanism 22 Second variable valve mechanism 23 Control shaft 24 First actuator 25 Drive shaft 26 Bearing bracket 27 Cylinder head 28 Swing cam 28a Cam body 28b Journal part 29 Valve stem 30 Valve lifter 31 Eccentric shaft portion 32 First link 33 Control eccentric shaft portion 34 Rocker arm 35 First connection pin 36 Second link 37 Second connection pin 38 Third connection pin 39 Cam surface 40 Sprocket 50 Eccentric sleeve bearing 51 Lever 51a Slit 52 Slide pin 53 Rod 54 Actuator piston 60 Cylinder 61 First hydraulic chamber 62 Second hydraulic chamber 63 First passage 64 Pump port 65 Hydraulic switching valve 66 Second passage 67 First on-off valve 68 Second on-off valve 69 First spring 70 Second spring 7 The first pilot passage 72 first cam 82 second cam 85 first motor 86 second pilot passage 73 spool valve 74 spring 80 piston spring 81 and the second motor

Claims (10)

A pressure chamber;
A pressure source for supplying fluid to the pressure chamber;
A displacement member that divides the pressure chamber into a first pressure chamber and a second pressure chamber, and moves in the pressure chamber in accordance with pressures of the first pressure chamber and the second pressure chamber;
A first passage communicating the pressure source and the first pressure chamber;
A second passage communicating the pressure source and the second pressure chamber;
A switching valve that is provided upstream of the first passage and the second passage and downstream of the pressure source, and controls fluid supply to the first pressure chamber and the second pressure chamber;
A first on-off valve provided in a first passage downstream of the switching valve and opening and closing the first passage;
A second on-off valve provided in a second passage downstream of the switching valve, for opening and closing the second passage;
With
The first on-off valve is closed when the pressure in the first pressure chamber is higher than the pressure in the second pressure chamber;
The second on-off valve is closed when the pressure in the second pressure chamber is higher than the pressure in the first pressure chamber;
A fluid pressure actuator that generates a driving force when the displacement member is moved by the fluid. A piston that reciprocates in the cylinder;
A first link coupled to the piston via a first coupling pin;
A second link that is rotatably mounted on a crankpin of the crankshaft and connected to the first link via a second connecting pin;
A third link coupled to the second link via a third coupling pin;
An eccentric shaft portion that is rotatably supported by the cylinder block and is eccentric with respect to the rotation shaft is detachably connected to the eccentric shaft portion, and is rotated according to the engine operating state to rotate the first link. An internal combustion engine variable compression ratio mechanism comprising a control shaft that variably controls the engine compression ratio by adjusting the position of the three links;
The fluid pressure actuator according to claim 1, wherein the control shaft is rotationally driven by the displacement member. The first on-off valve opens when the pressure in the second pressure chamber is equal to or higher than a predetermined value;
3. The fluid pressure actuator according to claim 1, wherein the second on-off valve opens when a pressure in the first pressure chamber is equal to or higher than a predetermined value. And a first small passage communicating the first on-off valve and the second pressure chamber;
A second small passage communicating the second on-off valve and the first pressure chamber;
One end of the first on-off valve is energized by an elastic body, and the other end is energized by the pressure of the second hydraulic chamber via the first small passage;
4. The fluid pressure actuator according to claim 3, wherein one end of the second on-off valve is biased by an elastic body, and the other end is biased by the pressure of the first hydraulic chamber via the second small passage.   The fluid pressure actuator according to claim 2, wherein the first on-off valve and the second on-off valve are driven by a rotating body synchronized with the crankshaft to open and close.   The fluid pressure actuator according to claim 2, wherein the first on-off valve and the second on-off valve are driven and opened / closed by an electric motor that is electrically synchronized with the crankshaft.   The fluid pressure actuator according to claim 6, wherein when the direction in which the control shaft is rotated and the direction of the torque applied to the control shaft are the same, both the first on-off valve and the second on-off valve are open.   The fluid pressure actuator according to any one of claims 2 to 7, wherein the pressure of the pressure source is smaller than a maximum pressure applied to the displacement member from the control shaft. In addition, a hydraulic variable valve mechanism that can change at least one of the operating angle of the intake valve and / or the exhaust valve and the central angle of the operating angle,
The fluid pressure actuator according to claim 2, wherein the pressure source is a hydraulic pressure source for driving the variable valve mechanism.   The fluid pressure actuator according to claim 9, wherein a pressure of the pressure source is smaller than a maximum pressure applied to the displacement member from the control shaft and larger than a pressure necessary for driving the variable valve mechanism.