VARIABLE DISPLACEMENT PUMP, OIL JET AND LUBLICATING SYSTEM USING VARIABLE DISPLACEMENT PUMP

Abstract

A variable displacement pump arranged to supply a hydraulic fluid to an oil jet arranged to inject the hydraulic fluid to a piston of an internal combustion engine, the variable displacement pump includes: a pump constituting section arranged to be driven and rotated by the internal combustion engine, and thereby to discharge the hydraulic fluid; a movable member arranged to decrease the flow rate discharged from the discharge portion by moving in one direction; a control section configured to move the movable member in the one direction by a predetermined amount when the discharge pressure becomes a first discharge pressure, and to further move the movable member in the one direction when the discharge pressure becomes a second discharge pressure larger than the first discharge pressure, the first discharge pressure being set smaller than the predetermined pressure at which the oil jet starts to inject the hydraulic fluid.

Description

BACKGROUND OF THE INVENTION

The present invention relates to a variable displacement pump used in an internal combustion engine and so on for a vehicle, an oil jet and a lubricating system using the variable displacement pump.

U.S. Patent Application Publication No. 2009-0101092 (corresponding to Japanese Patent Application Publication No. 2009-97424) discloses a conventional variable displacement pump including a first spring arranged to constantly act an urging force to a cam ring, and a second spring arranged to provide an urging force in a direction opposite to the urging force of the first spring when the cam ring is moved by a predetermined distance or more. In this conventional variable displacement pump, an eccentric state of the cam ring is varied in two stages (steps) by the relative urging forces of the springs, so that the discharge flow rate characteristic is varied in the two stages.

Moreover, this variable displacement pump is arranged to release a lock state of a valve timing control apparatus by the first discharge pressure before the cam ring is moved against the urging force of the first spring.

SUMMARY OF THE INVENTION

However, when this conventional variable displacement pump is also used for supplying the oil to an oil jet arranged to cool a piston of the internal combustion engine. When the oil in the first stage of the discharge pressure of the variable displacement pump, that is, the oil before the cam ring is moved is supplied, the unnecessary energy may be consumed until the discharge pressure to move the cam ring is obtained.

It is an object of the present invention to provide a variable displacement pump devised to suppress an energy consumption in an initial state of a discharge of an oil.

According to one aspect of the present invention, a variable displacement pump arranged to supply a hydraulic fluid to an oil jet arranged to inject the hydraulic fluid to a piston of an internal combustion engine when a pressure of the supplied hydraulic fluid becomes equal to or greater than a predetermined pressure, the variable displacement pump comprises: a pump constituting section arranged to be driven and rotated by the internal combustion engine, and thereby to discharge the hydraulic fluid entered from a suction portion to a plurality of operation chambers, from a discharge portion by volume variations of the operation chambers; a movable member arranged to decrease the flow rate of the hydraulic fluid discharged from the discharge portion by moving in one direction; and a control section configured to move the movable member in the one direction by a predetermined amount when the discharge pressure of the hydraulic fluid becomes a first discharge pressure, and to further move the movable member in the one direction when the discharge pressure of the hydraulic fluid becomes a second discharge pressure larger than the first discharge pressure, the first discharge pressure being set smaller than the predetermined pressure at which the oil jet starts to inject the hydraulic fluid.

According to another aspect of the invention, a lubricating system comprises: an oil jet arranged to inject a hydraulic fluid to a piston of an internal combustion engine when a pressure of a supplied hydraulic fluid becomes equal to or greater than a predetermined pressure; and a variable displacement pump arranged to supply the hydraulic fluid to the oil jet, the variable displacement pump including; a pump constituting section arranged to be driven and rotated by the internal combustion engine, and thereby to discharge the hydraulic fluid entered from a suction portion to a plurality of operation chambers, from a discharge portion by volume variations of the operation chambers; a movable member arranged to decrease the flow rate of the hydraulic fluid discharged from the discharge portion by moving in one direction; and a control section configured to move the movable member in the one direction by a predetermined amount when the discharge pressure of the hydraulic fluid becomes a first discharge pressure, and to further move the movable member in the one direction when the discharge pressure of the hydraulic fluid becomes a second discharge pressure larger than the first discharge pressure, the first discharge pressure being set smaller than the predetermined pressure at which the oil jet starts to inject the hydraulic fluid.

According to still another aspect of the invention, an oil jet comprises: a body including a hydraulic fluid supplying portion to which a hydraulic fluid is supplied, a hydraulic fluid introducing portion arranged to introduce the hydraulic fluid supplied to the hydraulic fluid supplying portion, and a valve seat formed between the hydraulic fluid supplying portion and the hydraulic fluid introducing portion; a valve element arranged to be seated on and released from the valve seat in accordance with the pressure of the hydraulic fluid supplied to the hydraulic fluid supplying portion, and thereby to open and close the hydraulic fluid supplying portion; an urging member arranged to urge the valve element in a valve closing direction in which the valve element closes the hydraulic fluid supplying portion, and to set a valve opening pressure of the valve element at which the valve element opens the hydraulic fluid supplying portion, to a value larger than the first discharge pressure; and an injection nozzle connected on a downstream side of the hydraulic pressure introducing portion, and arranged to inject the hydraulic fluid from an injection opening toward the piston, the hydraulic fluid supplying portion of body receiving the supply of the hydraulic fluid from a variable displacement pump including a pump constituting section arranged to be driven and rotated by an internal combustion engine, and thereby to discharge the hydraulic fluid entered from a suction portion to a plurality of operation chambers, from a discharge portion by volume variations of the operation chambers, a movable member arranged to decrease the flow rate of the hydraulic fluid discharged from the discharge portion by moving in one direction, and a control section configured to move the movable member in the one direction by a predetermined amount when the discharge pressure of the hydraulic fluid becomes a first discharge pressure, and to further move the movable member in the one direction when the discharge pressure of the hydraulic fluid becomes a second discharge pressure larger than the first discharge pressure, the valve element of the body being arranged to open the hydraulic fluid supplying portion to inject the hydraulic fluid to a piston of the internal combustion engine when a pressure of the supplied hydraulic fluid becomes equal to or greater than a predetermined pressure larger than the first discharge pressure.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is an exploded perspective view showing a variable displacement pump according to a first embodiment of the present invention.

FIG. 2 is a front view showing the variable displacement pump from which a cover member of the variable displacement pump is detached.

FIG. 3 is a sectional view taken along a section line A-A of FIG. 2.

FIG. 4 is a front view showing a pump housing of the variable displacement pump of FIG. 1.

FIG. 5 is a view for illustrating an operation of the variable displacement pump of FIG. 1.

FIG. 6 is a view for illustrating an operation of the variable displacement pump of FIG. 1.

FIG. 7 is a characteristic view showing a relationship between spring displacements and spring set loads of first and second coil springs.

FIG. 8 is a characteristic view showing a relationship between a discharge hydraulic pressure and an engine speed in a conventional variable displacement pump.

FIG. 9 is a longitudinal sectional view showing an internal combustion engine which employs an oil jet according to the first embodiment of the present invention.

FIG. 10 is a perspective view showing the oil jet according to the first embodiment of the present invention.

FIG. 11A is a longitudinal sectional view showing a valve closed state of the oil jet. FIG. 11B is a longitudinal sectional view showing a valve open state of the oil jet.

FIG. 12 is an overall sectional view showing a valve timing control apparatus according to the first embodiment of the present invention.

FIG. 13 is a sectional view showing the valve timing control apparatus in which a vane member is rotated to a most retarded position.

FIG. 14 is a sectional view showing the valve timing control apparatus in which the vane member is rotated in a most advanced side.

FIG. 15 is a sectional view showing a lock mechanism of the valve timing control apparatus of FIG. 12.

FIGS. 16A-16C are views showing operations of mechanisms at a stop of the engine. FIG. 16A is an illustrative view showing a state in which the vane member is controlled and rotated on the most retarded position. FIG. 16B is an illustrative view showing a state in which a lock piston is engaged in a lock hole. FIG. 16C is an illustrative view showing a state in which a spool valve element is held in a left side position.

FIGS. 17A-17C are views showing operations of the mechanism when an ignition key is switched to an ON state. FIG. 17a is an illustrative view showing a state in which the vane member is controlled and rotated to the most retarded position. FIG. 17B is an illustrative view showing a state in which the lock piston is pulled out from the lock hole. FIG. 17C is an illustrative view showing a state in which the spool valve element is held in the left side position.

FIGS. 18A-18C are views showing operations of the mechanisms when the engine is shifted to a middle engine speed. FIG. 18A is an illustrative view showing a state in which the vane member is controlled and rotated to the advance side. FIG. 18B is an illustrative view showing a state in which the lock piston is pulled out from the lock hole. FIG. 18C is an illustrative view showing a state in which the spool valve element is held in the right side position.

FIGS. 19A-19C are views showing operations of the mechanisms when the engine is in the idling operation. FIG. 19A is an illustrative view showing a state in which the vane member is controlled and rotated to the retard side. FIG. 19B is an illustrative view showing a state in which the lock piston is pulled out from the lock hole. FIG. 19C is an illustrative view showing a state in which the spool valve element is held in the left side position.

FIG. 20 is a front view showing a variable displacement pump according to a second embodiment of the present invention, from which a cover member is detached.

FIG. 21 is an exploded perspective view showing a variable displacement pump according to a third embodiment of the present invention.

FIG. 22 is a front view showing the variable displacement pump of FIG. 21 from which a cover member is detached, and in which an eccentric amount of a cam ring is maximized.

FIG. 23 is a front view showing the variable displacement pump of FIG. 21 from which the cover member is detached, and in which the eccentric amount of the cam ring is minimized.

FIG. 24 is a longitudinal sectional view showing the variable displacement pump of FIG. 21.

FIG. 25 is a front view showing an inside of a housing of the variable displacement pump of FIG. 21.

FIG. 26 is a longitudinal sectional view showing a state in which a solenoid valve is not energized in the variable displacement pump of FIG. 21.

FIG. 27 is a longitudinal sectional view showing a state in which the solenoid valve is energized in the variable displacement pump of FIG. 21.

FIG. 28 is a view showing a hydraulic circuit in the variable displacement pump of FIG. 21.

FIG. 29 is a graph showing a relationship between an engine speed and a hydraulic pressure in the variable displacement pump of FIG. 21.

FIG. 30 is a longitudinal sectional view showing a state in which a solenoid valve is not energized in a variable displacement pump according to a variation of the third embodiment of the present invention.

FIG. 31 is a longitudinal sectional view showing a state in which the solenoid valve is energized in the variable displacement pump according to the variation of the third embodiment of the present invention.

FIG. 32 is an exploded perspective view showing a variable displacement pump according to a fourth embodiment of the present invention.

FIG. 33 is a front view showing the variable displacement pump of FIG. 32 from which a cover member is detached, and in which an eccentric amount of the cam ring is maximized.

FIG. 34 is a front view showing the variable displacement pump of FIG. 33 from which the cover member is detached, and in which the eccentric amount of the cam ring is minimized.

FIG. 35 is a front view showing the cover member in the variable displacement pump of FIG. 32.

FIG. 36 is a back view showing the cover member in the variable displacement pump of FIG. 32.

FIG. 37 is a longitudinal sectional view showing a state in which a hydraulic directional control valve is not activated in a variable displacement pump according to a fifth embodiment of the present invention.

FIG. 38 is a longitudinal sectional view showing a state in which the hydraulic directional switching valve is activated in the variable displacement pump according to the fifth embodiment of the present invention.

FIG. 39 is a view showing a hydraulic circuit for the variable displacement pump according to the fifth embodiment of the present invention.

FIG. 40 is a graph showing a relationship between an engine speed and a hydraulic pressure in the variable displacement pump according to the fifth embodiment of the present invention.

DETAILED DESCRIPTION OF THE INVENTION

Hereinafter, variable displacement pumps according to embodiments of the present invention are illustrated with reference to drawings.

Variable displacement pump 01 according to embodiments are arranged to supply a lubricating oil to sliding portions of an internal combustion engine for a vehicle, to supply the lubricating oil through an oil jet to a piston, and to supply the lubricating oil to a valve timing control apparatus and a lock mechanism of the valve timing control apparatus.

First Embodiment

As shown in FIGS. 1-3, variable displacement vane pump 01 includes a pump housing 1 which is provided at a front end portion of a cylinder block of the internal combustion engine, and which has an opening of a one end closed by a cover member 2; a drive shaft 3 which penetrates a central portion within pump housing 1, and which is a driven and rotated by a crank shaft of the engine; a rotor 4 which is rotatably received within pump housing 1, and which has a central portion connected with drive shaft 3; a cam ring 5 which is a movable member swingably (pivotally) disposed radially outside rotor 4; and a pair of vane rings 6 and 6 each of which has a smaller diameter, and which is slidably disposed on both side surfaces of rotor 4 on the inner circumference side.

Pump housing 1 is integrally formed from an aluminum alloy. As shown in FIG. 4, pump housing 1 includes a bottom surface 1a which is a recessed shape, and on which an axial one end surface of cam ring 5 is arranged to be slid. Accordingly, bottom surface 1a is formed to have a flatness, a surface roughness and so on which have high accuracy. A sliding region (of bottom surface 1a) is formed by a machine working (machining).

Moreover, pump housing 1 includes a hole which is formed at a predetermined position on an inner circumference surface of pump housing 1, and into which a one end portion of a pivot pin 9 serving as a pivot point of cam ring 5 about which cam ring 5 is pivoted is inserted; and a pivot groove is which has a semi-circular cross section, which is formed at a predetermined position on an inner circumference surface of pump housing 1, and into which a one end portion of pivot pin 9 is inserted. Furthermore, pump housing 1 includes a seal surface 1s which has an arc recessed shape, and which is formed on the inner circumference surface of pump housing 1 on the left side of FIG. 2 at a position above a line M (hereinafter, referred to as a cam ring reference line) connecting a center of pivot pin 9 and a center of pump housing 1 (a center of drive shaft 3).

A seal member 14 (described later) provided to cam ring 5 is slid on seal surface 1s, so that seal surface 1s and also seal member 14 seal one end of a control hydraulic chamber 16 (described later) which is on upper end side in FIG. 2. As shown in FIG. 4, this seal surface 1s is formed into an arc surface having a predetermined radius.

As shown in FIG. 4, pump housing 1 includes a suction port 7 formed in bottom surface 1a on a left side of drive shaft 3, and a discharge port 8 formed in bottom surface 1a on a right side of drive shaft 3. Suction port 7 and discharge port 8 are disposed to confront each other.

As shown in FIG. 4, suction port 7 is connected with a suction opening 7a arranged to suck the oil within an oil pan (not shown). Discharge port 8 is connected from a discharge opening 8a through an oil main gallery (not shown) to the sliding portions of the engine, the valve timing control apparatus which is a variable valve actuating apparatus, and so on.

Suction port 7 includes an inside port portion 7b which has an arc shape, and an outside port portion 7c which has a substantially rectangular shape. Discharge port 8 includes an inside port portion 8b which has an arc shape, and an outside port portion 8c which is connected directly to discharge opening 8a.

Moreover, a bearing hole 1f for drive shaft 3 is formed at a substantially central portion of bottom surface 1a. The oil discharged from discharge port 8 is supplied to bearing hole 1f through a tip end recessed groove 10a of a feeding (supplying) groove 10 which has a small width, and which is formed into a substantially L-shape. Furthermore, the oil is supplied from an opening of feeding groove 10 to the both side surfaces of rotor 4, and side surfaces of vanes 11 (described later) so as to ensure lubricity.

As shown in FIGS. 1 and 3, cover member 2 is formed into a plate shape with a large thickness. Cover member 2 includes an inner side surface 2a which is a substantially flat surface. Cover member 2 includes a suction port 7′ and a discharge port 8′ formed in inner side surface 2a, like bottom surface 1a of pump housing 1. Suction port 7′ and discharge port 8′ are connected, respectively, with suction port 7 and discharge port 8. Cover member 2 includes a pin hole 2b which is formed at an end portion (on the right side in FIG. 3) of inner side surface 2a, and which receives the other end portion of pivot pin 9. Furthermore, cover member 2 includes a bearing hole 2c which is formed at a substantially central position of cover member 2, which penetrates cover member 2, through which drive shaft 3 is inserted, and which rotatably supports drive shaft 3.

This cover member 2 is positioned on pump housing 1 in the circumferential direction through a plurality of positioning pins 1P shown in FIG. 1, and mounted on pump housing 1 by a plurality of bolts B.

Drive shaft 3 is arranged to rotate rotor 4 in a clockwise direction of FIG. 2 by the rotational force transmitted from the crank shaft. A half on the left side of FIG. 2 around drive shaft 3 is a suction region. A half on the right side of FIG. 2 around drive shaft 3 is a discharge region.

As shown in FIGS. 1 and 2, rotor 4 includes seven slots 4a extending in the radial direction from the center side to the outside. Each of seven vanes 11 is slidably held in one of seven slots 4a, and arranged to be moved into and out of the one of seven slots 4a. Rotor 4 includes back pressure chambers 12 each of which is formed at a base end portion (radially inner end portion) of one of slots 4a, each of which has a substantially circular cross section, and each of which receives the discharge hydraulic fluid discharged into discharge port 8. Furthermore, rotor 4 includes recessed grooves 4b and 4b each of which has a circular (annular) recessed shape, which are formed on both end surfaces of rotor 4 in the axial direction, and which are arranged to hold vane rings 6 and 6 on the inner circumference side thereof so that vane rings 6 and 6 rotate in an eccentric state.

As shown in FIG. 2, each of vanes 11 includes an inside base end (radially inner end portion) which is slidably abutted on outer circumference surfaces of vane rings 6 and 6, and a tip end (radially outer end) which is slidably abutted on an inner circumference surface 5a of cam ring 5.

Moreover, a plurality of pump chambers 13 are liquid-tightly separated between adjacent two of vanes 11, inner circumference surface 5a of cam ring 5, the outer circumference surface of rotor 4, bottom surface 1a of pump housing 1, and inner side surface 2a of cover member 2. Each of pump chambers 13 is an operation chamber shaped like a sector. Each of vane rings 6 is arranged to push vanes 11 in the radially outward direction.

Cam ring 5 is integrally formed into a substantially cylindrical shape by an easily-worked sintered metal. Cam ring 5 includes a pivot raised portion 5b formed in a right side position of cam ring 5 in FIG. 2, on the cam ring reference line, on the outer circumference surface of cam ring 5. Cam ring 5 includes a pivot support groove 5k which has a semi-circular cross section, which extends in the axial direction, which is formed in a central position of the outer surface of pivot raised portion 5b, and which receives pivot pin 9 with pivot groove is so as to serve as an eccentric swing point about which cam ring 5 is swung.

Moreover, cam ring 5 includes a boss portion 5c which has a substantially inverse U-shape, and which is integrally formed with cam ring 5 at a position above cam ring reference line X, that is, an upper position on left side in FIG. 3. Cam ring 5 includes an arc surface 5d which is an arc raised portion, which is formed on an outside surface of boss portion 5c, and which confronts seal surface 1s. Furthermore, cam ring 5 includes a holding groove 5e which is a rectangular cross section, and which is formed in arc surface 5d. Seal member 14 is mounted and fixed in holding groove 5e. Seal member 14 seals the one end side of control hydraulic chamber 16. On the other hand, pivot support groove 5k of pivot raised portion 5b of cam ring 5 and pivot pin 9 seal the other end side of control hydraulic chamber 16. Arc surface 5d has a radius of curvature which is identical to that of seal surface 1s to form a minute constant clearance between arc surface 5d and seal surface 1s.

Seal member 14 is formed of, for example, a synthetic resin with a low abrasion resistance (low abrasion quality). Seal member 14 has an elongated shape extending in the axial direction of cam ring 5. Seal member 14 is pressed on seal surface 1s by a resilient (elastic) force of a resilient (elastic) member 15 made of rubber, and fixed on a bottom side of holding groove 5e. With this, good liquid-tightness of control hydraulic chamber 16 is constantly ensured.

As shown in FIGS. 1 and 3, cam ring 5 includes a pair of suction side cutout grooves 18a and 18b which are formed, in the circumferential direction, on the suction port 7's side, on both axial end surfaces of cam ring 5. The pair of suction side cutout grooves 18a and 18b introduce (flow) the oil in the suction region into pump chambers 13. Moreover, cam ring 5 includes discharge side cutout grooves 18c and 18d which are formed, in the circumferential direction, on the discharge port 8's side, on the both axial end surfaces of cam ring 5. Discharge side cutout grooves 18c and 18d introduce (flow) the oil within pump chambers 13 in the discharge region into discharge port 8.

Control hydraulic chamber 16 is separated in a substantially arc shape between the outer circumference surface of cam ring 5, pivot raised portion 5b and seal member 14. Control hydraulic chamber 16 is arranged to act the discharge hydraulic pressure introduced from discharge port 8, on a pressure receiving surface 5f in the outer circumference surface of cam ring 5 so that cam ring 5 is swung (pivoted) about pivot pin 9 in the counterclockwise direction of FIG. 2, and thereby to move cam ring 5 in a direction in which an eccentric amount (eccentricity) of cam ring 5 with respect to rotor 4 is decreased.

Cam ring 5 includes an arm 17 which is integrally formed with cam ring 5 on the outer circumference surface of the cylindrical body of cam ring 5 at a position opposite to pivot raised portion 5b, and which protrudes in the radially outward direction. As shown in FIGS. 1 and 2, this arm 17 includes an arm body 17a which has a rectangular plate shape, and which extends in the radial direction from the outer circumference surface of the cylindrical body of cam ring 5; and a raised portion 17b integrally formed on an upper surface of a tip end portion of arm body 17a.

Arm body 17a includes a protrusion 17c which has an arc curved shape, and which is integrally formed with arm body 17a on a lower surface of arm body 17a that is on a side opposite to raised portion 17b. On the other hand, raised portion 17b extends in a direction substantially perpendicular to arm body 17a. Raised portion 17b includes an upper surface 17d formed into a curved shape having a small radius of curvature.

A first spring receiving chamber 19 on a lower side of FIG. 2 and a second spring receiving chamber 21 on an upper side of FIG. 2 are formed at positions opposite to the position of pivot groove 1c of pump housing 1, that is, at upper and lower positions of arm 17. First spring receiving chamber 19 is formed coaxially with second spring receiving chamber 21.

First spring receiving chamber 19 has a substantially rectangular cross section. First spring receiving chamber 19 extends in the axial direction of pump housing 1. On the other hand, second spring receiving chamber 21 has a length shorter than a length of first spring receiving chamber 19. Like first spring receiving chamber 19, second spring receiving chamber 19 has a substantially rectangular cross section, and second spring receiving chamber 21 extends in the axial direction of pump housing 1.

As shown in FIG. 5, in second spring receiving chamber 21, there are provided a pair of retaining portions 23 and 23 which are formed into an elongated rectangular plate, which are formed integrally with pump housing 1 on inner side surfaces of second spring receiving chamber 21 at a lower opening portion 21a of second spring receiving chamber 21, and which extend in the inward directions to confront each other in widthwise direction of lower opening portion 21a. Raised portion 17b of arm 17 is arranged to be moved into or out of second spring receiving chamber 21 through opening portion 21a between retaining portions 23 and 23. Retaining portions 23 and 23 are arranged to restrict a maximum expansion of a second coil spring 22 described later.

A first coil spring 20 is received within first spring receiving chamber 19. First coil spring 20 is arranged to urge cam ring 5 through arm 17 in the clockwise direction of FIG. 2, that is, to urge cam ring 5 in a direction to increase the eccentric amount of the center of the inner circumference surface of cam ring 5 with respect to the center of the rotation of rotor 4.

First coil spring 20 is provided with a predetermined spring load W3. First coil spring 20 includes a lower end abutted on a bottom surface 19a of first spring receiving chamber 19, and an upper end constantly abutted on arc protrusion 17c on the lower surface of arm body 17a. First coil spring 20 is arranged to urge cam ring 5 in the direction to increase the eccentric amount of the center of the inner circumference surface of cam ring 5 with respect to the center of the rotation of rotor 4, that is, in the clockwise direction in FIG. 2.

A second coil spring 22 is received within second spring receiving chamber 21. Second coil spring 22 is arranged to urge cam ring 5 through arm 17 in the counterclockwise direction in FIG. 2.

This second coil spring 22 includes an upper end abutted on an upper wall surface 21b of spring receiving chamber 21, and a lower end abutted on raised portion 17b of arm 17 from the maximum eccentric position of cam ring 5 in the clockwise direction, to a position at which the lower end of this second coil spring 22 is retained by retaining portions 23 and 23 so as to provide the urging force in the counterclockwise direction of FIG. 2 to cam ring 5.

That is, second coil spring 22 is also provided with a predetermined spring set load in a direction to confront (opposite to) first coil spring 20. This spring set load of second coil spring 22 is set smaller than spring set load W3 of first coil spring 20. With this, cam ring 5 is set to an initial position (maximum eccentric position) by a spring load W1 which is a difference of the spring set loads between first coil spring 20 and second coil spring 22.

That is, first coil spring 20 and second coil spring 22 urges cam ring 5 through arm 17 in a direction in which cam ring 5 is constantly eccentric in the upward direction in a state in which spring load W1 is provided, that is, in a direction in which volumes of pump chambers 13 are increased. Spring load W1 is a load at which cam ring 5 starts to move when the hydraulic pressure is equal to or greater than a necessary hydraulic pressure P1 necessary for the valve timing control apparatus.

Second coil spring 22 is abutted on arm 17 when the eccentric amount of cam ring 5 between the center of the inner circumference surface of cam ring 5 and the center of the rotation of rotor 4 is equal to or greater than a predetermined amount. When the eccentric amount of cam ring 5 between the center of the inner circumference surface 5a of cam ring 5 and the center of the rotation of rotor 4 is smaller than the predetermined amount as shown in FIG. 5, second coil spring 22 is retained by retaining portions 23 and 23 to keep the compression state, and second coil spring 22 almost does not contact arm 17. A spring load W2 of first coil spring 20 at the swing movement amount of cam ring 5 at which the spring load of second coil spring 22 to arm 17 becomes zero by retaining second coil spring 22 by retaining portions 23 and 23 is a load at which cam ring 5 starts to move when the hydraulic pressure is a necessary hydraulic pressure P2 necessary for the piston oil jet and so on, or the necessary hydraulic pressure P3 at the maximum rotational speed of the crank shaft.

When cam ring 5 is pivoted in the clockwise direction by the spring force of first coil spring 20 as shown in FIG. 2, the upper surface of a join portion between arm body 17a and the cylindrical body is abutted on a lower surface of one of retaining portions 23 and 23, so that the further rotation of cam ring 5 in the clockwise direction is restricted. That is, the swing position of cam ring 5 is restricted to the initial set position (the maximum eccentric position) by the spring force of first coil spring 20.

Hereinafter, a basic operation of variable displacement pump 01 according to the first embodiment is illustrated. Moreover, a relationship between the control hydraulic pressure of the normal variable displacement pump, and the necessary hydraulic pressure necessary for the sliding portions of the engine, the valve timing control apparatus, and the cooling of the piston is illustrated.

When the valve timing control apparatus described later is used for improving the fuel consumption and for countermeasure for the exhaust air emission, the variable displacement pump is used as the operation source for the valve timing control apparatus. Accordingly, a high hydraulic pressure P1 shown by a broken line b of FIG. 8 is needed from the low engine speed so as to improve the operation responsiveness of the valve timing control apparatus.

Moreover, when oil jet 30 described later is used, high hydraulic pressure P2 is needed at the middle engine speed. A hydraulic pressure P3 is needed at the maximum engine speed, mainly for lubricating the bearing portion of the crank shaft. Therefore, the hydraulic pressure necessary for the entire of the internal combustion engine is a characteristic of the entire of the broken line connecting the broken lines b and c.

The relationship between middle engine speed necessary hydraulic pressure P2 and high engine speed necessary hydraulic pressure P3 is substantially P2<P3. Necessary hydraulic pressure P2 is often near necessary hydraulic pressure P3. Accordingly, it is desirable that the hydraulic pressure is not increased even when the engine speed is increased from the middle engine speed to the high engine speed in a region (D).

In this example, as shown by a solid line of FIG. 8, the pump discharge pressure does not reach hydraulic pressure P1 from the start of the internal combustion engine to the low engine speed including the idling. Accordingly, arm body 17a of arm 17 of cam ring 5 is abutted on the one of retaining portions 23 of housing 1 by the difference between the spring loads of first coil spring 20 and second coil spring 22, so that cam ring 5 is in the activation stop state (cf. FIG. 2).

At this time, the eccentric amount of cam ring 5 is maximized, and the pump capacity is maximized. The discharge hydraulic pressure is suddenly increased in accordance with the increase of the engine speed. Accordingly, the discharge hydraulic pressure becomes a characteristic shown by (A) on the solid line of FIG. 8.

Then, when the pump discharge hydraulic pressure is increased in accordance with the increase of the engine speed and reaches hydraulic pressure Pf shown in FIG. 8, the hydraulic pressure introduced into control hydraulic chamber 16 is increased. With this, cam ring 5 starts to compress first coil spring 20 acted to arm 17, and cam ring 5 is pivoted about pivot pin 9 in the counterclockwise direction to be eccentric. The hydraulic pressure Pf is a first hydraulic pressure which is a cam activation pressure for activating (starting the movement of) cam ring 5. The hydraulic pressure Pf is set to be sufficiently greater than necessary hydraulic pressure P1 of the valve timing control apparatus.

With this, the pump capacity is decreased. Accordingly, the increase characteristic of the discharge hydraulic pressure is decreased as shown in a region (B) of FIG. 8. Then, as shown in FIG. 5, second coil spring 22 is retained by retaining portions 23 and 23 to keep the compression state. Cam ring 5 is swung in the counterclockwise direction to a state in which the load of second coil spring 22 is not acted to upper surface 17d of arm raised portion 17b.

In a state shown in FIG. 5, the load of second coil spring 22 is not acted to cam ring 5 from this instant. Cam ring 5 can not be swung and becomes a retained state until the discharge hydraulic pressure reaches the hydraulic pressure P2 (hydraulic pressure P2 within control hydraulic chamber 16), and becomes greater than spring load W2 of first coil spring 20. Accordingly, the discharge hydraulic pressure becomes an increasing characteristic shown by (C) of FIG. 8 in accordance with the increase of the engine speed. However, the discharge hydraulic pressure does not become sudden increase shown by (A) of FIG. 8 since the pump capacity is decreased by the decrease of the eccentric amount of cam ring 5.

Moreover, when the discharge pressure increases equal to or greater than hydraulic pressure Ps (P2) by the increase of the engine speed, cam ring 5 is swung against the spring force of spring load W2 of first coil spring 20 through arm 17 to compress first coil spring 20, as shown in FIG. 6. Accordingly, the pump capacity is further decreased in accordance with the swing movement of cam ring 5, so that the increase of the discharge hydraulic pressure becomes small. Then, the engine speed reaches the maximum engine speed while the discharge hydraulic pressure keeps the state of the characteristic shown by (D) of FIG. 8.

Accordingly, the discharge pressure (solid line) at the high rotational speed of the pump sufficiently approaches the necessary (requested) hydraulic pressure (broken line). Therefore, it is possible to effectively suppress the power loss.

FIG. 7 shows a relationship between displacements of first and second coil springs 20 and 22 or the swing angle of cam ring 5, and spring loads W1 and W2. That is, the spring force of set load W1 of first coil spring 20 is provided to cam ring 5 at the initial state from the start of the internal combustion engine to the low engine speed. Accordingly, cam ring 5 can not be moved until the discharge pressure exceeds spring load W1. After the discharge pressure exceeds spring load W1, first coil spring 20 is compressed, so that the load of first coil spring 20 is increased. On the other hand, second coil spring 22 approaches a free length so that the load of second coil spring 22 is decreased. Consequently, the spring load is increased. This inclination is a spring constant.

At a position of cam ring 5 in FIG. 5, the load becomes spring load W2 of first coil spring 20. The spring load W2 of first coil spring 20 increases in a discontinuous manner. After the discharge hydraulic pressure exceeds spring load W2, first coil spring 20 is compressed, so that the load (of first coil spring 20) is increased. However, only one coil spring is acted. Therefore, the spring constant is decreased, so that the inclination is varied.

As described above, when the discharge hydraulic pressure reaches hydraulic pressure P1 by the increase of the internal combustion engine speed, cam ring 5 starts to move to suppress the increase of the discharge hydraulic pressure. When cam ring 5 is moved by a predetermined movement amount in the counterclockwise direction as shown in FIG. 5, the spring force of second coil spring 22 is eliminated, and the spring constant becomes small. Moreover, the spring load W2 of first coil spring 20 is increased in the discontinuous manner. Accordingly, the swing movement of cam ring 5 starts again after the discharge hydraulic pressure is increased to the hydraulic pressure P2 (cf. FIG. 6). That is, the relative spring load of first and second coil springs 20 and 22 are acted, and the spring characteristic is non-linear state. With this, cam ring 5 acts a special swing movement.

In this way, the characteristic of the discharge hydraulic pressure becomes characteristics shown by (A)-(D) of FIG. 8 by the non-linear characteristic of the spring forces of coil springs 20 and 22. Accordingly, the control hydraulic pressure (solid line) can sufficiently approach the necessary (requested) hydraulic pressure (broken line). Therefore, it is possible to sufficiently decrease the power loss caused by the unnecessary increase of the hydraulic pressure.

Next, an oil jet 30 according to the first embodiment of the present invention is illustrated.

As shown in FIG. 9, oil jet 30 is provided to an internal combustion engine 31. In the internal combustion engine 31, a crank shaft 34 is rotatably supported by a bearing (not shown) within a crank chamber 33 separated by a crank case of a cylinder block 32. Moreover, a piston 36 is slidably disposed within a cylindrical cylinder wall 37 formed at an upper portion of the crank case. Piston 36 is connected through a con rod 35 to crank shaft 34.

Within a wall of cylinder wall 37, there is formed a water jacket 37a in which the cooling water is circulated. Within a partition wall 38 between the crank case and cylinder wall 37, there is formed a main oil gallery 39 arranged to supply the oil (the lubricating oil) discharged from variable displacement pump 01 to the sliding portions of the engine.

Within the lower portion of partition wall 38, there is formed a connection passage 38a which is connected with main oil gallery 39, and which extends in the upward and downward directions, as shown in FIGS. 11A and 11B. In the lower portion of connection passage 38a, there is formed a mounting hole including an internal-thread portion 38b formed on an inner circumference surface of the mounting hole.

Oil jet 30 is attached (mounted) at the lower portion of partition wall 38. Oil jet 30 is arranged to inject the oil for the lubrication and the cooling, to a portion between the inner circumference surface of cylinder wall 37 and piston 36.

As shown in FIGS. 10, 11A and 11B, this oil jet 30 includes a cylindrical holding member 40 made of an aluminum alloy; a valve body 41 formed into a cylindrical shape, and inserted from the below into an insertion hole 40a formed within holding member 40; a protruding portion 42 which is integrally formed in the outside portion of holding member 40, and which serves as a positioning member; and a nozzle 43 formed on the outside portion of holding member 40 at a position opposite to protruding portion 42.

Holding member 40 includes an annular passage portion 44 formed between insertion hole 40a and an outer circumference surface of valve body 41; and a mounting groove 40b which is formed in the outside portion of holding member 40, and in which a base end portion 43a of nozzle 43 is mounted and fixed.

Valve body 41 is formed of an iron metal such as sintered alloy. Valve body 41 includes an external thread portion 41a which is formed on an outer circumference surface of the upper portion of valve body 41, and which is screwed into internal thread portion 38b. Moreover, valve body 41 includes an oil supply hole 45 that is a hydraulic fluid supplying portion which is formed within the upper end portion of valve body 41, which extends in the axial direction, and which is connected with connection passage 38a; an oil introduction hole 47 which is formed on a lower side of oil supply hole 45, which is connected with oil supply hole 45 in a continuous manner, and which is arranged to movably hold a ball valve element 46; and a seat surface 45a which is formed into an annular shape, which is formed in a stepped portion between oil supply hole 45 and oil introduction hole 47, and on which valve ball element 46 is arranged to be seated.

Furthermore, valve body 41 includes a plurality of radial holes 48 formed along the diameter direction in a circumferential wall of the lower end portion of valve body 41. Radial holes 48 connect oil introduction hole 47 and passage portion 44. Valve body 41 includes a flange portion 41b integrally formed with the outer circumference of the lower portion of valve body 41. This flange portion 41 is arranged to press and fix base end portion 43a of nozzle 43 and also holding member 40 on partition wall 38 when valve body 41 is screwed and fixed in partition wall 38 through external thread portion 41a and internal thread portion 38b.

Protruding portion 42 is mounted in a positioning hole 38c formed in partition wall 38 when holding member 40 is fixed in partition wall 38 through valve body 41, so as to position holding member 40, and to prevent the rotation of holding member 40.

Nozzle 43 is raised (extends) in an inclined state from a base end portion 43a located on the holding member 40's side to a tip end portion 43b. Nozzle 43 is disposed so that tip end portion 43b is located in a lower portion within cylinder wall 37. Nozzle 43 includes an elongated hydraulic hole 43c which extends within nozzle 43 in the axial direction, and which has a one end portion opened to passage portion 44; and a nozzle portion 43d formed at the tip end portion of hydraulic hole 43c, and arranged to direct the lower portion of piston 36.

A valve spring 50 is held by a plug-shaped retainer 49 fit in the lower end portion of oil introduction hole 47 by the press fit. Valve spring 50 is an urging member having a coil shape. Ball valve element 46 is urged by the urging force (the spring force) of valve spring 50 in a direction in which ball valve element 46 is seated on seat surface 45a, that is, in a direction to close the opening of the lower end of oil supply hole 45.

The spring load of valve spring 50 (that is, the valve opening pressure of ball valve element 46) is set to pressure P2 (cf. FIG. 8) which is sufficiently larger than first discharge pressure Pf of variable displacement pump 01, and which is slightly smaller than second discharge pressure Ps that is a working (operating) pressure of cam ring 5.

Hereinafter, an operation of oil jet 30 is illustrated. First, drive shaft 3 of variable displacement pump 01 is rotated in response to the start of the engine, so that the pressurized oil is supplied to main oil gallery 39 to lubricate the sliding portions of the engine. In the initial state of the start of the engine, the pump discharge pressure is first discharge pressure Pf, as shown in FIG. 8. Accordingly, ball valve element 46 is seated on seat surface 45a by the spring force of valve spring 50 to keep the valve closed state, as shown in FIG. 11A.

Then, when the pump discharge pressure is increased and the hydraulic pressure within oil supply hole 45 becomes equal to or greater than the hydraulic pressure P2, that is, the spring load of valve spring 50, valve spring 50 is compressed to open ball valve element 46, as shown in FIG. 11B. With this, oil supply hole 45 is connected with oil introduction hole 47, the oil supplied from main oil gallery 39 through connection passage 38a to oil supply hole 45 flows (enters) from oil introduction hole 47 through radial holes 48 into connection passage 44. Moreover, the oil flows through hydraulic hole 43c of nozzle 43, and the oil is injected from nozzle portion 43d into the inside from the lower direction of piston 36, as shown in FIG. 9.

In this way, the oil discharged from variable displacement pump 01 is not injected from oil jet 30 to piston 36 until the discharge pressure of the oil becomes equal to or greater than first discharge pressure Pf, and reaches the hydraulic pressure P2 slightly smaller than second discharge pressure Ps. Therefore, it is possible to effectively suppress the energy loss in the initial stage of the pump discharge of variable displacement pump 01.

Moreover, as described above, first discharge pressure Pf is set smaller than the valve opening pressure of ball valve element 46 of oil jet 30. Accordingly, oil jet 30 does not inject the oil in the engine speed region (normal region) which is used at the normal running of the vehicle. Therefore, the oil supply amount to the sliding portions of the engine is increased while the excessive oil discharge amount of the pump is suppressed. Therefore, it is possible to decrease the friction of the pump and the internal combustion engine, and to improve the fuel consumption.

Moreover, at the cold state of internal combustion engine 31, it is possible to suppress the injection of the low temperature oil by oil jet 30 to piston 36. Therefore, it is possible to improve the warm-up characteristic, and to decrease the exhaust emission.

The structure of oil jet 30 is not limited to the structure of the above-described embodiment. Holding member 40 may be formed integrally with nozzle 43. Nozzle 43 may be fixed to valve body 41 by the brazing. Moreover, a plunger may be employed as the valve element, in place of the ball.

Next, the valve timing control apparatus is illustrated below.

This valve timing control apparatus is applied to the intake side. As shown in FIGS. 12-15, this valve timing control apparatus includes a timing sprocket 51 which is a driving rotational member driven and rotated through a timing chain by the crank shaft (not shown) of the engine; a cam shaft 52 arranged to rotate relative to timing sprocket 51; a vane member 53 which is a driven rotational member fixed at an end portion of cam shaft 52, and rotatably received within timing sprocket 51, and a hydraulic pressure supply and discharge mechanism 54 arranged to rotate vane member 53 in a normal direction or in a reverse direction by the hydraulic pressure.

Timing sprocket 51 includes a housing 55 having a teeth portion 55a integrally formed on an outer circumference of housing 55, and engaged with the timing chain, and which rotatably receives vane member 53; a front cover 56 closing an opening of a front end of housing 55; and a rear cover 57 closing an opening of the rear end of housing 55. These housing 55, front cover 56 and rear cover 57 are integrally fixed by four small diameter bolts 58 from the axial direction of the cam shaft.

Housing 55 has a cylindrical shape having front and rear both ends each having an opening. Housing 55 includes four partition wall portions 60 which are shoes that are arranged on the inner circumference surface at regular intervals of 90 degrees in the circumferential direction, and that protrudes in the radially inward direction. Each of partition wall portions 60 has a substantially trapezoid cross section. Each of partition wall portions 60 extends in the axial direction of housing 55. Each of partition wall portions 60 has both axial end surfaces which are same planes with both end surfaces of housing 55. Moreover, each of four partition wall portions 60 includes a bolt insertion hole 61 which is located at a substantially central position of the each of partition wall portions 60, and into which one of bolts 58 is inserted. Each of partition wall portions 60 includes an inner end surface (radially inner surface) formed into an arc shape to correspond to an outer circumference of vane rotor 64 (described later) of vane member 53, and a holding groove which is formed on the inner end surface, and which extends in the axial direction. A U-shaped seal member 62 and a plate spring (not shown) arranged to press seal member 62 in the inward direction are fit in and held by the holding groove of the each of partition wall portions 60.

Front cover 56 includes a bolt insertion hole 56a which has a relatively large diameter, and which is formed at a substantially central portion of front cover 56; and four bolt holes which are formed in the outer circumference portion of front cover 56, and each of which is connected with one of bolt insertion holes 61 of housing 55.

Rear cover 57 includes a bearing hole 57a which is formed at a substantially central portion of rear cover 57, and which rotatably supports a front end portion 52a of cam shaft 52; and four internal thread holes which are formed in an outer circumference portion of rear cover 57, and into which one of bolts 58 is screwed.

Cam shaft 52 is rotatably supported by an upper end portion of the cylinder head through a cam bearing (not shown). Cam shaft 52 includes a cam which is integrally formed on the outer circumference surface of cam shaft 52 at a predetermined position, and which is arranged to open an intake valve (not shown) through a valve lifter.

Vane member 53 is integrally formed by the sintered alloy. Vane member 53 includes an annular (circular) vane rotor 64 located at a central portion, and fixed at a front end portion of cam shaft 52 by cam bolt 63; four vanes 65 integrally formed with vane rotor 64, and arranged on the outer circumference surface of vane rotor 64 at intervals of 90 degrees in the circumferential direction. Vane rotor 64 includes an axial hole 64a which is located at a substantially central position of vane rotor 64, and into which cam bolt 63 is inserted; and a mounting groove 64b in which front end portion 52a of cam shaft 52 is inserted and mounted. Vane rotor 64 is fixed on front end portion 52a of cam shaft 52 by cam bolt 63 from the axial direction.

One of four vanes 65 has a substantially trapezoid shape having a large circumference width in the substantially circumferential direction. Each of the other three of the four vanes 65 has an elongated rectangular shape. These four vanes 65 are disposed in the circumferential direction at predetermined angular positions to attain the weight balance of the entire of vane member 53. Moreover, each of vanes 65 is disposed between adjacent two of partition wall portions 60. Each of vanes 65 has a holding groove formed at a central portion of the outer circumference surface. A U-shaped seal member 66 and a plate spring 66a are fit and mounted in each of the holding grooves. Seal member 66 is slidably abutted on the inner circumference surface of housing 55. Plate spring 66a is arranged to push seal member 66 toward the inner circumference surface of housing 55.

Moreover, an advance fluid pressure chamber 67 and a retard fluid pressure chamber 68 are formed, respectively, on both sides of each vane 65. Accordingly, four advance fluid pressure chambers 67 and four retard fluid pressure chambers 68 are separated between vanes 65 and partition wall portions 60.

As shown in FIG. 12, hydraulic pressure supply and discharge mechanism 54 includes two hydraulic passage systems including a first hydraulic passage 69 arranged to supply and discharge the hydraulic pressure of the lubricating oil, to and from advance fluid pressure chambers 67; a second hydraulic passage 70 arranged to supply and discharge the hydraulic pressure to and from retard fluid pressure chambers 68. First and second hydraulic passages 69 and 70 are connected through a flow passage switching valve 73 to a supply passage 71 and a drain passage 72 which are main oil galleries for supplying the lubricating oil for the engine. One-way variable displacement pump 01 arranged to pressurize the hydraulic fluid within oil pan 74, and to supply this pressurized hydraulic fluid is provided to supply passage 71. Moreover, a lower end of drain passage 72 is connected with oil pan 74.

As shown in FIGS. 12 and 13, first hydraulic passage 69 is formed between flow passage switching valve 73 and each of advance fluid pressure chambers 67. First hydraulic passage 69 includes a first passage portion 69a formed from the inside of the cylinder head to the inside of the cam bearing and cam shaft 52 in the axial direction; and four bifurcated passages 69b which are formed by bifurcating within vane rotor 64 in the substantially radial directions from the grooves on the front end side of cam shaft 52, and each of which connects first passage portion 68a and one of advance fluid pressure chambers 67.

On the other hand, second hydraulic passage 70 is formed between flow passage switching valve 73 and each of retard fluid pressure chambers 68. Second hydraulic passage 70 includes a second passage portion 70a formed from the inside of the cylinder head to the insides of the cam bearing and cam shaft 52 in the axial direction; and four second bifurcated passages 70b which are formed by bifurcating in the radial direction from the radial hole of cam shaft front end portion 52a to the inside of vane rotor 64, and each of which connects second passage portion 70a and one of retard fluid pressure chambers 68.

A phase varying mechanism is constituted by vane member 53, housing 55, advance fluid pressure chambers 67, retard fluid pressure chambers 68, and hydraulic pressure supply and discharge mechanism 54.

As shown in FIG. 12, flow passage switching valve 73 is a solenoid valve which is 4-port and 2-position type. Flow passage switching valve 73 includes a valve body 77 which has a cylindrical shape with a bottom, and which is fixed within valve hole 76 formed within the cylinder head; a solenoid 78 integrally fixed to one end portion of valve body 77; and a spool valve element 79 slidably disposed within valve body 77.

Valve body 77 includes a supply port 80 located at a substantially central position in the axial direction, and arranged to connect supply passage 71 and the inside of valve body 77; and first and second ports 81 and 82 which are located on both sides of supply port 80 in the axial direction, which are arranged to connect, respectively, the end portions of first hydraulic passage 69 and second hydraulic passage 70, and the inside of valve body 77, and which extend in the radial direction. Moreover, valve body 77 includes first and second drain ports 83 and 84 which are formed, respectively, on both sides of first and second ports 81 and 82, and which connect, respectively, the inside of valve body 77 and drain passage 72.

Solenoid 78 includes an electromagnetic coil 78b provided within a solenoid casing 78a; a fix core 78c arranged to be excited by an energization to electromagnetic coil 78b; a movable plunger 78d arranged to be slid by the excitation of fix core 78c, and thereby to push and move spool valve element 79. Electromagnetic coil 78b is connected through a harness (not shown) to an electronic controller 86.

Spool valve element 79 includes a first land portion 79a located at a substantially central portion of spool valve element 79, and arranged to open and close supply port 80 in accordance with the sliding position of spool valve element 79 in the axial direction; and second and third land portions 79b and 79c disposed on both sides of first land portion 79a in the axial direction, and arranged to relatively open and close first and second ports 81 and 82 and drain ports 83 and 84. Moreover, this spool valve element 79 is urged to a maximum left position, that is, a position to connect supply port 80 and second port 82, and to connect first port 81 and drain port 83, by a spring force of a return spring 85 mounted between a spring retainer 77a provided on the other end side of valve body 77, and an outer end surface of third land portion 79c. Furthermore, spool valve element 79 is arranged to be controlled to move against the spring force of return spring 85, to a maximum right position or a predetermined central position, by a control current from electronic controller 86.

Electronic controller 86 is configured to sense a current driving state by signals from a crank angle sensor (not shown) arranged to sense the engine speed, and an air flow meter arranged to sense an intake air amount, and various sensors such as a throttle opening sensor and a water temperature sensor arranged to sense a water temperature of the engine.

This electronic controller 86 is configured to perform a switching control of the flow passages by applying or breaking (cutting off) a pulse control current to electromagnetic coil 78a of flow passage switching valve 73, in accordance with the driving state of the engine.

Moreover, between vane 65 with the maximum width and housing 55, there is provided a lock mechanism 87 arranged to restrict the rotation of vane member 53 with respect to housing 55, or to release the restriction of the rotation of vane member 53.

As shown in FIGS. 12 and 15, this lock mechanism 87 includes a sliding hole 88 formed between vane 65 with the large width and rear cover 57, and formed within vane 65 in the axial direction of cam shaft 52; a lock piston 89 which is a cylindrical shape with a cover, and which is slidably provided within sliding hole 88; a lock hole 90a which is formed in an engagement hole forming section 90 that has a cup-shaped cross section, and that is fixed in a fixing hole formed in rear cover 57, and which is an abutment portion which a taper tip end portion 89a of lock piston 89 is engaged with or disengaged from; a coil spring 92 which is held by a spring retainer 91 fixed on the bottom surface side of sliding hole 88, and which is a third urging member arranged to urge lock piston 89 toward lock hole 90a.

Lock piston 89 includes a large diameter flange 89b which is integrally formed with the outer circumference on the rear end side of lock piston 89, and which is arranged to receive the pressure; and a tip end portion 89a arranged to be engaged with lock hole 90a by the spring force of coil spring 92 at a position at which vane member 53 is rotated to the most retarded position, and thereby to lock the relative rotation between timing sprocket 51 and cam shaft 52.

As shown in FIG. 15, lock piston 89 is moved in a backward direction against the spring force of coil spring 92 by the hydraulic pressure supplied from advance fluid pressure chambers 67 into lock hole 90a through a first hydraulic hole 93a formed in vane member 53, or by the hydraulic pressure supplied from retard fluid pressure chambers 68 into a pressure receiving chamber 89c between the large diameter flange portion 89a and the stepped portion of sliding hole 88, through a second hydraulic hole 93b formed in vane member 53, so as to release the engagement with lock hole 90a.

Coil spring 92 serves as a lock state holding mechanism to hold the lock state between vane member 53 and housing 55. Coil spring 92 has a spring force which is set to a value by which the air accumulated in retard fluid pressure chambers 68 at the start of the engine is not largely compressed and deformed by the pressure compressed by the pressurized hydraulic pressure supplied from variable displacement pump 01, and which is set to a value by which the air accumulated in retard fluid pressure chambers 68 is compressed and deformed when the discharged hydraulic pressure reaches the hydraulic pressure Px in the initial state (A) shown in FIG. 8.

Hereinafter, the operation of the valve timing control apparatus is illustrated with reference to FIGS. 16-19. The operation of variable displacement pump 01 is stopped at the stop of the engine, so that the supply of the hydraulic pressure to advance fluid pressure chambers 68 and retard fluid pressure chambers 68 is stopped. As shown in FIGS. 13 and 16A, vane member 53 is rotated in a direction opposite to the rotational direction of cam shaft 52 (shown by an arrow) to the most retarded side, by the alternating torque generated in advance in cam shaft 52 immediately after the stop of the engine.

At this instant, tip end portion 89a of lock piston 89 of lock mechanism 87 is engaged with lock hole 90a by the spring force of coil spring 92 as shown in FIG. 16B, so as to restrict the free rotation of vane member 53.

Moreover, the energization from electronic controller 86 to flow passage switching valve 73 is cut off (shut off). Accordingly, spool valve element 79 is urged to the maximum left side position by the spring force of return spring 85, as shown in FIG. 16C.

Next, when the ignition key is switched to the ON state to start the engine, the control current from electronic controller 86 is not outputted to electromagnetic coil 78b for a few seconds from the start of the cranking. Accordingly, spool valve element 79 is urged to the maximum left side position by the spring force of return spring 85, as shown in FIG. 17C. Therefore, supply port 80 and second port 82 are connected with each other, and second land portion 79b closes second drain port 84. At the same time, first and third land portions 79a and 79c connect first port 81 and first drain port 83.

Accordingly, the hydraulic pressure (discharge pressure) discharged from variable displacement pump 01 flows from supply passage 71 through supply port 80 into valve body 77, as shown by arrows in FIG. 17C. Then, this hydraulic pressure flows from second port 82 directly to second hydraulic passage 70. This hydraulic pressure flows through second bifurcated passage 70b into retard fluid pressure chambers 68.

Accordingly, as shown in FIG. 17A, vane member 53 is held, by the low hydraulic pressure supplied into retard fluid pressure chambers 68, to a state in which vane member 53 is positioned in the most retarded side. Consequently, it is possible to improve the performance of the start of the engine.

In this case, the air accumulated in retard fluid pressure chambers 68 is pressurized by the low hydraulic pressure, so that the air accumulated in retard fluid pressure chambers 68 pushes vane member 53 to the most retarded side with the low hydraulic pressure.

On the other hand, when the internal pressure of retard fluid pressure chambers 68 is increased, this hydraulic pressure is supplied from second hydraulic hole 93b to pressure receiving chambers 89c, and acted to the pressure receiving surface of large diameter flange 89b. With this, as shown in FIG. 17B, lock piston 89 is moved in the backward direction against the spring force of coil spring 92, and pulled out from lock hole 90. Consequently, vane member 53 is released from the lock state to allow the free rotation. However, vane member 53 is held to the maximum retarded position like the stop of the engine since the hydraulic pressure within retard fluid pressure chambers 68 is high.

This timing at which end portion 89a of lock piston 89 is pulled out from lock hole 90a is a timing at which the discharge hydraulic pressure characteristic of variable displacement pump 01 becomes discharge pressure Px which is lower than first discharge pressure Pf, and which is at the sudden increase before the compression of first coil spring 20 in the region (A) of FIG. 8, that is, a timing at which two or three seconds elapsed from the switching of the ignition key to the ON state.

Then, when the engine speed becomes, for example, the middle engine speed region after the start of the cranking, electronic controller 86 energizes electromagnetic coil 78b of flow passage switching valve 73 so as to excite fix core 78c. With this, spool valve element 79 is moved in the right direction from the position shown in FIG. 17C through movable plunger 78d to the maximum right side position shown in FIG. 18C. Consequently, spool valve element 79 closes the connection between first port 81 and first drain port 83, and connects supply port 80 and first port 81. At the same time, spool valve element 79 connects second port 82 and second drain port 84.

Accordingly, as shown in FIG. 18C, the discharge hydraulic pressure of variable displacement pump 01 flows from supply passage 71 into supply port 80 and valve body 77. Then, this discharge hydraulic pressure flows from first port 81 into first passage portion 69a of first hydraulic passage 69. This discharge hydraulic pressure is supplied through bifurcated passage 69b into advance fluid pressure chambers 67 to increase the pressure in the advance fluid pressure chambers 67. On the other hand, the hydraulic fluid within retard fluid pressure chambers 68 is returned from second hydraulic passage 70 and so on through second drain port 84 to oil pan 74 so as to decrease the pressure within retard fluid pressure chambers 68.

Accordingly, in lock piston 89, the hydraulic pressure of pressure receiving chamber 89c is lowered. However, as shown in FIG. 18B, lock piston 89 is held to a state in which lock piston 89 is pulled out from lock hole 90a against the spring force of coil spring 92 by the high hydraulic pressure supplied into lock hole 90a from first hydraulic hole 93a in accordance with the increase of the hydraulic pressure of advance fluid pressure chambers 67. With this, as shown in FIG. 18A, vane member 53 is rotated by the high hydraulic pressure of advance fluid pressure chambers 67 from the position shown in FIG. 13 in the rightward direction, that is, in a direction identical to the rotational direction of cam shaft 52, so that the relative rotational phase between the crank shaft and cam shaft 52 is rapidly varied to the advanced side.

Accordingly, the valve overlap between the intake valve and the exhaust valve is slightly increased. Therefore, it is possible to decrease the discharge amount of HC in the exhaust gas by the effect of the internal EGR, as described later.

Moreover, when the engine is shifted, for example, to the high engine speed region, the energization from electronic controller 86 to electromagnetic coil 78b is held, the hydraulic pressure is constantly supplied to advance fluid pressure chambers 67. Accordingly, vane member 53 is further rotated in the same direction, and held in the maximum rotation position as shown in FIG. 14. With this, the relative rotational phase between the crank shaft and cam shaft 52 is varied to the most advanced position. Consequently, the valve overlap becomes large, and it is possible to improve the output (power) of the engine.

Moreover, when the operation of the engine is shifted to the idling operation, the control current from electronic controller 86 to electromagnetic coil 78b is cut off. Accordingly, as shown in FIG. 19C, spool valve element 79 is moved in the maximum left direction by the spring force of return spring 85. Consequently, supply port 80 and second port 82 are connected with each other, and second land portion 79b closes second drain port 84. At the same time, first and third land portions 79a and 79c connect first port 81 and first drain port 83.

Accordingly, the hydraulic pressure discharged from variable displacement pump 01 flows from supply passage 71 through supply port 80 into valve body 77, as shown by an arrow of FIG. 19C. This hydraulic pressure flows from second port 82 directly to second hydraulic passage 70. Then, the hydraulic pressure is supplied through second bifurcated passages 70b into retard fluid pressure chambers 68. On the other hand, the hydraulic pressure of advance fluid pressure chambers 67 is discharged from first hydraulic passage 69 through first port 81, first drain port 83, and drain passage 72 into oil pan 74, so that advance fluid pressure chambers 67 are brought into the low pressure state.

In this case, lock piston 89 is held to a pulled-out state in which lock piston 89 is pulled out from lock hole 90e, by the hydraulic pressure in pressure receiving chamber 89c which receives the high pressure within retard fluid pressure chambers 68, as shown in FIG. 19B. Accordingly, the free rotation of vane member 53 is allowed. Consequently, this vane member 53 is pivoted in the most retarded side by the high hydraulic pressure supplied into the retard angle chambers 68, as shown in FIG. 19A. With this, it is possible to improve the combustion, and to improve the stability of the idle operation of the engine.

As described above, in this example, it is possible to improve the responsiveness of the operation of the valve timing control apparatus at the start of the engine by the special structures by using first and second coil springs 20 and 22 of variable displacement pump 01.

That is, variable displacement pump 01 supplies the lubricating oil discharged from the discharge opening through discharge port 8, to the sliding parts of the engine. Moreover, variable displacement pump 01 is also used as the source of the operation of the valve timing control apparatus. In variable displacement pump 01, it is possible to improve the initial increase of the initial discharge hydraulic pressure (region (A)) shown in FIG. 8, as described above. Accordingly, it is possible to improve, for example, the responsiveness of the relative rotational phase between timing sprocket 55 and cam shaft 52 to the retarded side immediately after the start of the engine.

Second Embodiment

FIG. 20 shows a variable displacement pump according to a second embodiment of the present invention. The variable displacement pump according to the second embodiment has a basic structure substantially identical to the basic structure such as the pump constituting section of the variable displacement pump according to the first embodiment. However, in the second embodiment, there are provided two control hydraulic chambers 16 which are arranged to push cam ring 5 to increase the eccentric amount, and which are located on upper and lower sides of pivot pin 9.

That is, the control hydraulic chamber in the first embodiment represents first hydraulic chamber 16a. Moreover, in this example, pump housing 1 includes a recessed groove 24 which is formed on the lower side of pivot pin 9, and which has a substantially L-shape. This recessed groove 24 constitutes (forms) a second control hydraulic chamber 16b. Furthermore, in the lower portion of recessed groove 24, there is formed a second seal surface 24a. This second seal surface 24a is formed into an arc shape around a center of pivot pin 9.

On the other hand, cam ring 5 includes a raised portion 25 which is formed integrally with cam ring 5 at a portion to confront recessed groove 24, and which has a substantially triangular shape. Raised portion 25 includes a second arc surface 25a which is located in a portion to confront second seal surface 24a, and which has an arc shape around the center of pivot pin 9. Second arc surface 25a includes a holding groove which is located at a tip end portion of second arc surface 25a, and which has a substantially rectangular cross section. The holding groove of second arc surface 25a receives a seal member 26 slidably abutted on seal surface 24a, and a resilient (elastic) member 27 which has a rectangular cross section, and which pushes seal member 26 toward second seal surface 24a.

Seal surface 24a has a length of the arc by which seal member 26 can be slidably abutted on seal surface 24a when the eccentric amount of cam ring 5 with respect to the center of rotor 4 is swung from the maximum eccentric amount shown in FIG. 2 to the minimum eccentric amount shown in FIG. 14.

Second control hydraulic chamber 16b is connected with discharge port 8 through a connection groove 1g formed in bottom surface 1a of pump housing 1. Accordingly, the discharge pressure is acted on second pressure receiving surface 5g of cam ring 5 on the outer circumference to confront second control hydraulic chamber 16b, like first pressure receiving surface 5f receiving the discharge pressure of first control hydraulic chamber 16a.

Second arc surface 25a has a radius of curvature smaller than a radius of curvature of first arc surface 5d on first seal member 14's side. Second pressure receiving surface 5g has a surface area smaller than a surface area of first pressure receiving surface 5f. Accordingly, when the discharge pressures of first and second control hydraulic chambers 16a and 16b are acted, respectively, on pressure receiving surfaces 5f and 5g, a swing torque in the counterclockwise direction of FIG. 20 is generated in cam ring 5, like the first embodiment. However, the hydraulic torque from second control hydraulic chamber 16b acted only on second pressure receiving surface 5g is in the clockwise direction. Accordingly, a part of the hydraulic torque generated in cam ring 5 is canceled. Therefore, the swing torque of cam ring 5 is smaller than that of cam ring 5 in the first embodiment when the discharge pressures are identical to each other.

Accordingly, the spring forces of coil springs 20 and 22 can be set to the smaller values. Consequently, it is possible to decrease radii of coil springs 20 and 22, and thereby to decrease the entire size of the vane pump.

Third Embodiment

FIGS. 21-29 show a variable displacement pump according to a third embodiment of the present invention. The variable displacement pump 01 according to the third embodiment is substantially identical to the variable displacement pump 01 according to the first and second embodiments. Accordingly, repetitive illustrations are omitted.

That is, this variable displacement pump 01 includes a pump housing 111 which has a U-shaped cross section, and which includes a pump receiving chamber 113; a cover member 112 which closes an opening of an one end side of pump housing 111; a drive shaft 114 which penetrates a substantially central portion of pump chamber 113, and which is driven and rotated by the crank shaft of the engine; a rotor 115 which is rotatably received within pump receiving chamber 113, and which includes a central portion connected with drive shaft 114; seven vanes 116 each of which is moved into and out of one of slots 115a that are formed in an outer circumference portion of rotor 115, and that extend in the radial directions; a cam ring 117 which is disposed within pump housing 111, and which is arranged to be swung to be eccentric with respect to a center of the rotation of rotor 115; a single coil spring 118 which is received within pump housing 111, and which is an urging member arranged to constantly urge cam ring 117 in a direction to increase the eccentric amount (eccentricity) of cam ring 117 with respect to the center of the rotation of rotor 115; and vane rings 119 and 119 slidably disposed on the inner circumference portion of the both axial side surfaces of rotor 115. A pump constituting (forming) section is constituted (formed) by drive shaft 114, rotor 115, vanes 116 and cam ring 117.

As shown in FIGS. 24 and 25, pump housing 111 includes a bearing hole 111a which is formed at a substantially central portion of a bottom surface 113a of pump receiving chamber 113, which rotatably supports one end portion of drive shaft 114, and which penetrates bottom surface 113a (pump housing 111). Moreover, as shown in FIG. 25, pump housing 111 includes a support groove 111b which has a semi-circular cross section, which is formed at a predetermined position of an inner circumference wall of pump receiving chamber 113 that is an inside surface of pump housing 111, and which swingably supports cam ring 117.

Moreover, on the inner circumference wall of pump receiving chamber 113, there are formed first and second sliding surfaces 111c and 111d located on both sides of a cam ring reference line M connecting a center of bearing hole 111a and a center of support groove 111b to sandwich cam ring reference line M, and on which seal members 130 and 130 (described later) disposed on the outer circumference surface of cam ring 117 are slidably abutted. These seal sliding surfaces 111c and 111d have, respectively, arc surfaces which are formed about the center of support groove 111b, and which have predetermined radii R1 and R2. These seal sliding surfaces 111c and 111d have, respectively, circumferential lengths by which seal members 130 and 130 can be constantly slidably abutted on these seal sliding surfaces 111c and 111d in the eccentric swing region of cam ring 117. Accordingly, cam ring 117 is slid on and introduced by seal sliding surfaces 111c and 111d when cam ring 117 is swung to be eccentric. Consequently, it is possible to obtain smooth activation (eccentric swing movement) of cam ring 117.

As shown in FIGS. 22 and 25, there are formed a suction port 121 which is formed on bottom surface 113a of pump receiving chamber 113 in the outer circumference region of bearing hole 111a, and which is a suction portion that has a substantially arc recessed shape, and that is opened in a region (suction region) in which the inside volumes of pump chambers 120 are increased in accordance with the pump operation; and a discharge port 122 which is formed on bottom surface 113a of pump receiving chamber 113 in the outer circumference region of bearing hole 111a, and which is a discharge portion that is a substantially arc recessed shape, and that is opened in a region (discharge region) in which the inside volumes of pump chambers 120 are decreased in accordance with the pump operation. Suction port 121 and discharge port 122 are positioned at positions to sandwich bearing hole 111a, and to confront each other.

Suction port 121 is connected with an introduction passage 124 extending from a substantially central position of suction port 121 toward spring receiving chamber 128. In introduction passage 124, there is formed a suction hole 121a which penetrates a bottom wall of pump housing 111, and which is opened to the outside. With this, as shown in FIG. 28, the lubricating oil stored in oil pan 152 of the engine is sucked through suction hole 121a and suction port 121 to pump chambers 120 in the suction region, based on the negative pressure generated in accordance with the pump operation of the pump constituting member.

Suction hole 121a and also introduction passage 124 confront the outer circumference region of cam ring 117 on the pump suction side. Suction hole 121a is arranged to introduce the suction pressure to the outer circumference region of cam ring 117 on the pump suction side. With this, the pressure of the outer circumference region of cam ring 117 on the pump suction side which is adjacent to pump chambers 120 in the suction region becomes the suction pressure or the atmospheric pressure. Accordingly, it is possible to suppress the leakage of the lubricating oil from pump chambers 120 in the suction region to the outer circumference region of cam ring 117 on the pump suction side. In this case, the pump suction side is a region on a left side of a cam ring eccentric direction line N (described later) in FIG. 22.

Discharge port 122 is connected with an introduction passage 125 extending from a start end portion of discharge port 122 to confront a first control hydraulic chamber 131 (described later) defined on the outer circumference side of cam ring 117. At a terminal end portion of introduction passage 125, there is formed a discharge hole 122a which penetrates the bottom wall of pump housing 111, and which is opened to the outside.

This discharge hole 122a is connected through main oil gallery 39 to the sliding portions of the engine, the valve timing control apparatus, and oil jet 30.

By the thus-constructed configuration, the lubricating oil which is pressurized by the pump operation of the pump constituting section, and which is discharged from pump chambers 120 in the discharge region is supplied through discharge port 122 and discharge hole 122a to the sliding portions within the engine and the valve timing control apparatus.

Discharge hole 122a and also introduction passage 125 confront the outer circumference region of cam ring 117 on the pump discharge side. Discharge hole 122a is arranged to introduce the discharge pressure to the outer circumference region of cam ring 117 on the pump discharge side. In this case, the pump discharge side represents a region on a right side of the cam ring eccentric direction lien N (described later) in FIG. 22.

At a portion near the start end portion of discharge port 122, there is formed a connection groove 123 connecting discharge port 122 and bearing hole 111a. The lubricating oil is supplied through connection groove 123 to bearing hole 111a. Moreover, the lubricating oil is supplied to the side portions of rotor 115 and vanes 116 to ensure the lubricity of the sliding portions.

This connection groove 123 is formed so as not to correspond to the direction in which each of the vanes 116 is moved into and out of one of slots 115a. With this, vanes 116 are suppressed to drop into connection groove 123 when vanes 116 are moved into and out of slots 115a.

Cover member 112 is formed into a substantially plate shape. Cover member 112 has a portion which is located on the outer side surface (on the left side in FIG. 24), which corresponds to bearing hole 111a of pump housing 111, and which has a slightly large thickness. Cover member 112 includes a bearing hole 112a which is formed in this large thickness portion, which penetrates the large thickness portion, and which rotatably supports the other end portion of drive shaft 114. This cover member 112 has a substantially flat inner side surface (on the right side in FIG. 24). This cover member 112 is mounted on the open end surface of pump housing 111 by a plurality of bolts 126.

Drive shaft 114 is arranged to rotate rotor 115 in the clockwise direction of FIG. 22 by the rotational force transmitted from the crank shaft. A left half of FIG. 22 with respect to a line N (hereinafter, referred to as a cam ring eccentric direction line) perpendicular to cam ring reference line M at the center of drive shaft 114 is the pump suction side. A right half of FIG. 22 with respect to cam ring eccentric direction line N is the pump discharge side.

As shown in FIGS. 21 and 22, rotor 115 includes the plurality of slots 115a formed to extend in the radially outward directions from the center side; a plurality of back pressure chambers 115b each of which has a substantially circular cross section, each of which is formed at a radially inside end (inside base end) of one of slots 115a, and which is arranged to receive the discharge fluid discharged to discharge port 122. With this, vanes 116 are pushed in the radially outward directions by the centrifugal force caused by the rotation of rotor 115, and the hydraulic pressures of back pressure chambers 115b.

Each of vanes 116 includes a tip end portion (radially outside end) slidably abutted on an inner circumference surface of cam ring 117, and a base end portion (radially inside end) slidably abutted on an outer circumference surfaces of vane rings 119 and 119. With this, pump chambers 120 are liquid-tightly separated by the outer circumference surface of rotor 115, the inner side surfaces of adjacent two of vanes 116 and 116, the inner circumference surface of cam ring 117, bottom surface 113a of pump receiving chamber 113 of pump housing 111, and the inner side surface of cover member 112, even when the engine speed is low and the centrifugal force and the hydraulic pressures of back pressure chambers 115b are small.

Cam ring 117 is integrally formed into a substantially cylindrical shape by the sintered metal. Cam ring 117 includes a pivot portion 117a which is formed into a substantially arc raised shape, which is mounted in support groove 111b of pump housing 111, and which serves as an eccentric swing support portion about which cam ring 117 is swung; and an arm portion 117b which is located at a position opposite to pivot portion 117a with respect to the center of cam ring 117, which is liked (connected) with coil spring 118, and which extends in the axial direction to protrude.

Within pump housing 111, there is formed a spring receiving chamber 128 located at a position opposite to support groove 111b, and connected with pump receiving chamber 113 through connection portion 127 having a predetermined width L. Coil spring 118 is received in this spring receiving chamber 128.

This coil spring 118 is resiliently held between the bottom surface of spring receiving chamber 128 and the lower surface of the tip end portion of arm portion 117b extending through connection portion 127 to spring receiving chamber 128. Coil spring 118 has a predetermined set load W. Arm portion 117b includes a support protrusion 117i which is formed into a substantially arc shape, which is formed on a lower surface of the tip end portion of arm portion 117b, and which is engaged with the inner circumference side of coil spring 118. One end portion of coil spring 118 is supported by support protrusion 117i.

Coil spring 118 constantly urges cam ring 117 through arm portion 117b in a direction in which the eccentric amount of cam ring 117 is increased (in the clockwise direction of FIG. 22) by the resilient force based on set load W. With this, cam ring 117 is brought to a state in which the upper surface of arm portion 117b is pressed by the urging force of coil spring 118 on a stopper portion 128a protruding on a cover portion of spring receiving chamber 128, in the non-activation state of cam ring 117 shown in FIG. 22. Cam ring 117 is restricted at the position at which the eccentric amount of cam ring 117 is maximized.

In this way, arm portion 117b extends in a direction opposite to pivot portion 117a. The tip end portion of arm portion 117b is urged by coil spring 118. With this, the maximum torque can be generated in cam ring 117. Accordingly, it is possible to decrease the size of coil spring 118, and thereby to decrease the size of the pump itself.

Moreover, cam ring 117 includes a pair of first and second seal constituting portions 117c and 117d which are formed on the outer circumference portion of cam ring 117, which have substantially triangular cross section, which extend in the axial direction to protrude, and which include first and second seal surfaces 117g and 117h that have arc surfaces concentric with seal sliding surfaces 111c and 111d, and that confront first and second seal sliding surfaces 111c and 111d. First and second seal surfaces 117g and 117h of first and second seal constituting portions 117c and 117d include first and second holding grooves 117e and 117f having a substantially rectangular cross section, and extending in the axial direction. Seal holding grooves 117e and 117f receive, respectively, seal members 130 and 130 slidably abutted on seal sliding surfaces 111c and 111d at the eccentric swing movement of cam ring 117.

Seal surfaces 117g and 117h are formed about the center of pivot portion 117a. Seal surfaces 117g and 117h have predetermined radii R3 and R4 slightly smaller than radii R1 and R2 of the corresponding seal sliding surfaces 111c and 111d. Between seal surfaces 117g and 117h and seal sliding surfaces 111c and 111d, there are formed, respectively, minute clearances C.

Seal members 130 and 130 are made of, for example, fluorine resin with a low frictional characteristic. Each of seal members 130 and 130 has an elongated shape extending linearly in the axial direction of cam ring 117. Seal members 130 and 130 are pressed on seal sliding surfaces 111c and 111d by the resilient forces of resilient members 129 and 129 made from rubber, and disposed on the bottom portions of seal holding grooves 117e and 117f. With this, it is possible to constantly ensure the good liquid-tightness of pressure chambers 131 and 132 described later.

In the outer circumference region of cam ring 117 on the pivot portion 117a's side of cam ring eccentric direction line N which is the pump discharge side (on the right side in FIG. 22) in the non-activation state of cam ring 117, there are separated a first control hydraulic chamber 131 and a second control hydraulic chamber 132 defined on both sides of pivot portion 117a to sandwich pivot portion 117a, by the outer circumference surface of cam ring 117, pivot portion 117a, seal members 130 and 130, and the inside surface of pump housing 111.

In this example, the entire of first and second control hydraulic chambers 131 and 132 are set within the region on the pump discharge side, on the outer circumference region of cam ring 117. It is preferable to set the entire of first and second control hydraulic chambers 131 and 132 in a region overlapped with the discharge region which is the pressurized region in the radial direction, that is, a region which confront pump chambers 120 that become constantly the positive pressure to sandwich the circumference wall of cam ring 117.

The discharge pressure discharged to discharge port 122 is constantly introduced through introduction passage 125 to first control hydraulic chamber 131. The discharge pressure is acted to a first pressure receiving surface 133 which is constituted by the outer circumference surface of cam ring 117 that confronts first control hydraulic chamber 131, and which receives the force acted to counteract (block) the urging force of coil spring 118. With this, cam ring 117 receives the swing force (movement force) in a direction (the counterclockwise direction) to decrease the eccentric amount of cam ring 117.

That is, this first control hydraulic chamber 131 is arranged to urge cam ring 117 through first pressure receiving surface 133 in a direction in which the center of cam ring 117 approaches (become concentric to) the center of the rotation of rotor 115. With this, first control hydraulic chamber 131 serves for the movement amount control in the concentric direction of cam ring 117.

On the other hand, second control hydraulic chamber 132 is arranged to receive the discharge pressure through an introduction hole 135 which penetrates the bottom wall of pump housing 111, and which is connected with discharge hole 122a through a solenoid valve 140 (described later) which is controlled in accordance with the driving state of the engine. Accordingly, the discharge pressure is acted to second pressure receiving surface 134 which is constituted by the outer circumference surface of cam ring 117 that confronts second hydraulic chamber 132, and which receives the force acted in a direction to assist the urging force of coil spring 118. With this, cam ring 117 receives the swing force in a direction (in the clockwise direction of FIG. 22) in which the eccentric amount of cam ring 117 is increased.

As shown in FIG. 22, second pressure receiving surface 134 has a pressure receiving area S2 smaller than a pressure receiving area S1 of first pressure receiving area 133. The urging force in the eccentric direction of cam ring 117 by the urging force based on the internal pressure of second control hydraulic chamber 132 and the urging force of coil spring 118, and the urging force by first control hydraulic chamber 131 are balanced to keep a predetermined force relationship. The urging force by second control hydraulic chamber 132 assists the urging force of coil spring 118.

Second control hydraulic chamber 132 is arranged to act the discharge pressure supplied through solenoid valve 140 to second pressure receiving surface 134, and thereby to assist the urging force of coil spring 118. With this, second control hydraulic chamber 132 serves for the movement amount control of cam ring 117 in the eccentric direction.

As shown in FIG. 28, this variable displacement pump 01 is provided with solenoid valve 140 which is a component different from variable displacement pump 01, and which is operated in accordance with the driving state of the engine based on the excitation current from ECU 151 mounted on the vehicle. Discharge hole 122a and introduction hole 135 are connected with each other through this solenoid valve 140. With this, first control hydraulic chamber 131 and second control hydraulic chamber 132 are connected with each other in the open state of solenoid valve 140.

As shown in FIGS. 26 and 27, this solenoid valve 140 includes a valve body 141 which has a cylindrical shape, and which includes an one end (on the right side in FIGS. 26 and 27) opened, and the other end (on the left side in FIGS. 26 and 27) closed; a valve element 142 which is disposed within valve body 141 to be slid in the axial direction, and which includes first and second land portions 142a and 142b located on both side end portions of valve element 142, and arranged to be slid on an inner circumference surface of valve body 141; a spring 143 which is received in a back pressure chamber 145 separated on the other end side of valve body 141 by second land portion 142b of valve element 142, and which is arranged to urge valve element 142 toward the one end side of valve body 141; an electromagnetic unit 144 mounted on the opening of valve body 143, and which is arranged to move a rod 144b in accordance with the energization, and thereby to move valve element 142 toward the other end side of valve body 141 in the axial direction against the urging force of spring 143.

Valve body 141 includes an IN port 141a which is formed in the circumference wall, which penetrates the circumference wall, and which is connected with discharge hole 122a; an OUT port 141b which is formed in the circumference wall, which penetrates the circumference wall, and which is connected with introduction hole 135; a drain port 141c which is formed in the circumference wall, which penetrates the circumference wall, and which is connected with suction port 121 or the outside. Moreover, valve body 141 includes a back pressure port 141d which is formed in a side wall of the other end, which penetrates the side wall of the other end, which is connected with suction port 121 or the outside, and which is constantly opened to back pressure chambers 145.

Valve element 142 includes a substantially central portion in the axial direction which has a smaller diameter, and land portions 142a and 142b which define an annular space 146 between the central portion of valve element 142 and valve body 141. OUT port 141b, and IN port 141a or drain port 141c are connected through annular space 146.

Electromagnetic unit 144 has a conventional structure. Electromagnetic unit 144 includes a coil unit 144a having a bobbin around which a coil is wound, and a yoke mounted on the thus-constructed bobbin; an armature (not shown) which is made of a magnetic material, which is disposed radially inside coil unit 144a, and which is moved into and out of coil unit 144a in the axial direction; and a rod 144b which is connected with the armature, and which is moved into and out of (in the forward and rearward directions) with the armature in accordance with the energization state.

As shown in FIG. 26, solenoid valve 140 is a normally open type. When the excitation current is applied to coil unit 144a, IN port 141a and OUT port 141b are connected with each other through annular space 146. With this, the discharge pressure is introduced into second control hydraulic chamber 132. In this case, drain port 141c is opened to back pressure chamber 145.

On the other hand, when the excitation current is applied to coil unit 144a, valve element 142 is pressed and returned toward the other end side of valve body 141 against the urging force of spring 143 by the pressing force of rod 144b, as shown in FIG. 27. With this, IN port 141a is closed by first land portion 142a of valve element 142. OUT port 141b is connected through annular space 146 with drain port 141c. With this, second control hydraulic chamber 132 is opened to the suction pressure or the atmospheric pressure.

By the thus-constructed structure, in variable displacement pump 01, a relationship of relative forces acted to cam ring 117 between the internal pressure of first control hydraulic chamber 131, the urging force of coil spring 118, and the internal pressure of second control hydraulic chamber 132 that is controlled by solenoid valve 140 is controlled so as to control the eccentric amount of cam ring 117. With this, the variation of the inside volumes of pump chambers 120 are controlled at the pump operation by controlling the eccentric amount, so that the discharge pressure characteristic of variable displacement pump 01 is controlled.

Hereinafter, the operation of variable displacement pump 01 according to the third embodiment, that is, the discharge pressure control of the pump based on the eccentric amount control of cam ring 117 is illustrated with reference to FIGS. 22, 23 and 29.

When the valve timing control apparatus is activated, the requested (necessary) pressure of the discharge pressure of variable displacement pump 01 becomes hydraulic pressure P1 in FIG. 29, as described above. That is, the valve timing control apparatus is set to activate (start) by the low hydraulic pressure P1 immediately after the start of the engine.

The requested hydraulic pressure of the crank metal at the high engine speed is a hydraulic pressure P2 in FIG. 29 at the low load or the low hydraulic oil temperature. On the other hand, the requested hydraulic pressure of the crank metal at the high engine speed is a hydraulic pressure P4 in FIG. 29 at the high load or the high hydraulic oil temperature.

Moreover, at the high load of the engine, oil jet 30 is used for cooling the piston. The valve opening pressure of ball valve element 46 of this oil jet 30 is set to a hydraulic pressure P3 in FIG. 29 at a predetermined engine speed n which is the middle engine speed.

At the low load or the low hydraulic oil temperature, variable displacement pump 01 is set to a low pressure characteristic X which is a first discharge pressure characteristic that satisfies one of hydraulic pressure P1 and hydraulic pressure P2 of FIG. 29, or both of the hydraulic pressure P1 and hydraulic pressure P2. At the high load or the high hydraulic oil temperature, variable displacement pump 01 is set to a high pressure characteristic Y which is a second discharge pressure characteristic that satisfies one of the hydraulic pressure P3 and the hydraulic pressure P4, or both of hydraulic pressure P3 and hydraulic pressure P4.

The operation characteristic of cam ring 117, that is, first and second operation hydraulic pressures Px and Py which are needed for the activation of cam ring 117 are varied by switching of ON/OFF states of solenoid valve 140. The appropriate hydraulic pressure characteristic is selected from both hydraulic pressure characteristics X and Y in accordance with the driving state of the engine, so as to satisfy the requested hydraulic pressure of the engine.

In this example, as shown in FIG. 29, low pressure characteristic X is set to the hydraulic pressure characteristic shown by a broken line connecting requested hydraulic pressure P1 of the valve timing control apparatus and requested hydraulic pressure P2 at the high engine speed at the low load or the low hydraulic fluid temperature. On the other hand, high pressure characteristic Y is set to the hydraulic pressure characteristic shown by a solid line connecting requested hydraulic pressure P3 which is the valve opening pressure of oil jet 30 at the middle engine speed at the high load or the high hydraulic fluid temperature, and requested hydraulic pressure P4 at the high engine speed at the high load or the high hydraulic fluid temperature.

That is, in variable displacement pump 01, spring load W of coil spring 118 is set to first activation hydraulic pressure Px. IN port 141a is closed (shut off) by applying the excitation current from ECU 151 to solenoid valve 140 at the low load or the low oil temperature. With this, the discharge pressure is introduced only to first control hydraulic chamber 131.

Accordingly, the eccentric amount of cam ring 117 is held in the maximum state until the internal pressure of the first control hydraulic chamber 131 reaches first activation hydraulic pressure Px (cf. FIG. 22). The discharge pressure is suddenly increased in accordance with the increase of the engine speed.

When the internal pressure of first control hydraulic chamber 131 reaches first activation hydraulic pressure Px by the increase of the discharge pressure, cam ring 117 is swung about pivot portion 117a in the downward direction of the cam ring eccentric direction line N, that is, in a direction in which the eccentric amount is decreased (cf. FIG. 23). Consequently, the variations of the volumes of pump chambers 120 are decreased at the operation of the pump. Therefore, the increase of the discharge pressure according to the increase of the engine speed becomes gentle, so that low pressure characteristic X shown in FIG. 29 can be obtained.

On the other hand, when it is shifted from the low load or the low oil temperature state to the high load or the high oil temperature state, the excitation current from ECU 151 to solenoid valve 140 is shut off. With this, IN port 141a and OUT port 141b are connected with each other. Accordingly, the discharge pressure is introduced to first control hydraulic chamber 131 and also second control hydraulic chamber 132.

The pressure acted on second pressure receiving surface 134 of second control hydraulic chamber 132 serves to assist the urging force of coil spring 118. Accordingly, cam ring 117 is not activated even when the internal pressure of first control hydraulic chamber 131 reaches first activation hydraulic pressure Px of FIG. 29. Cam ring 117 is held in a state in which the eccentric amount of cam ring 117 is maximized, until the difference of the hydraulic pressures which are acted on first pressure receiving surface 133 and second pressure receiving surface 134 by the internal pressure of first control hydraulic chamber 131 and the internal pressure of second control hydraulic chamber 132 reaches the urging force of coil spring 118 (cf. FIG. 22).

As shown in FIG. 29, the eccentric amount of cam ring 117 is held in the maximum state at the high load or the high oil temperature state, until the discharge pressure reaches second activation hydraulic pressure Py so that the difference of the hydraulic pressures acted on first pressure receiving surface 133 and second pressure receiving surfaces 134 by the internal pressure of first control hydraulic chamber 131 and the internal pressure of second control hydraulic chamber 132 becomes equal to the urging force of coil spring 118. Accordingly, the discharge pressure is largely increased in accordance with the increase of the engine speed.

Then, cam ring 117 is swung in a direction in which the eccentric amount is decreased when the internal pressure of first control hydraulic chamber 131 reaches second activation hydraulic pressure Py (cf. FIG. 23). With this, the variations of the volumes of pump chambers 120 become small at the pump operation, and the increase of the discharge pressure according to the increase of the engine speed becomes gentle. Consequently, high pressure characteristic Y can be obtained as shown in FIG. 29.

In variable displacement pump 01, in principle, the pump discharge pressure characteristic is shifted to high pressure characteristic Y when ECU 151 judges that the high pressure is needed by the engine speed, the load, the oil temperature and so on.

Normally, the pump discharge pressure characteristic is shifted to high pressure characteristic Y when the load of the engine, the oil temperature and so on are high. In the above-described illustration, the high pressure characteristic Y is used at the high load of the engine or the high oil temperature. However, for example, the valve timing control apparatus may need the hydraulic pressure higher than requested hydraulic pressure P1. In this case, ECU 151 switches solenoid valve 140 in accordance with the activation signal of the valve timing control apparatus. Even at the low load of the engine, the low oil temperature or so on, the pump discharge pressure characteristic is shifted to high pressure characteristic Y.

That is, in this example, requested hydraulic pressure P1 is set to the normal requested hydraulic pressure of the valve timing control apparatus. Requested hydraulic pressure P1 is set to minimum requested hydraulic pressure in the valve timing control apparatus, in accordance with the specification and so on of the vehicle employing this valve timing control apparatus.

Moreover, when it is shifted again from the high load or the high oil temperature state to the low load or the low oil temperature state, ECU 151 applies the excitation current to solenoid valve 140 again, so that solenoid valve 140 becomes the energization state shown in FIG. 27. With this, second control hydraulic chamber 132 is opened to the atmospheric pressure or the suction pressure. Accordingly, the activation of cam ring 117 depends on the relationship of the force between the internal pressure of first control hydraulic chamber 131 and the urging force of coil spring 118. With this, the discharge pressure characteristic of the pump is varied to low pressure characteristic X. Consequently, it is possible to decrease the discharge pressure which is not needed for shifting to the low load or the low oil temperature state, and to suppress the power loss of the engine.

In this variable displacement pump 01, ECU 151 switches solenoid valve 140 in accordance with the driving information such as the engine speed, the load of the engine, and the oil temperature, so that the activation characteristic of cam ring 117 is varied. Accordingly, it is possible to select the discharge pressure characteristic suitable for the engine speed, the load of the engine, the oil temperature and so on. With this, it is possible to cut waste of the work of the pump, and to suppress (minimize) the power loss of the engine.

Moreover, in this variable displacement pump 01, the operation control of cam ring 117 does not need the complex control such as the duty control. Furthermore, variable displacement pump 01 does not need a high-precision work (processing) of the shape of the port, the tuning of the valve opening characteristic and so on of solenoid valve 140. Accordingly, it is possible to readily attain the operation control of cam ring 117 by the simple control by the switching of ON/OFF of solenoid valve 140, and by the simple structure using the general solenoid valve 140. Therefore, it is possible to decrease the manufacturing cost of the pump.

In variable displacement pump 01, the internal pressures of pump chambers 120 in the discharge region are acted on the inner circumference surface of cam ring 117 on pivot portion 117a's side, as shown by a bold solid arrow of FIG. 23. Accordingly, cam ring 117 is pressed in the right direction of FIG. 23 along the cam ring reference line M, that is, toward support groove 111b, so that pivot portion 117a is pressed against support groove 111b.

However, in variable displacement pump 01 of this example, control hydraulic chambers 131 and 132 are disposed radially outside cam ring 117 on the pump discharge side, that is, so as to confront these pump chambers 120 in the discharge region to sandwich the circumference wall of cam ring 117. As shown by bold broken arrows in FIG. 23, the internal pressures of both of control hydraulic chambers 131 and 132 are acted, respectively, on cam ring 117 so as to push cam ring 117 in a direction opposite to support groove 111b. Consequently, it is possible to decrease the tight abutment (engagement) of pivot portion 117a against support groove 111b. Therefore, it is possible to decrease the friction between pivot portion 117b and support groove 111b at the eccentric swing movement of cam ring 117.

Accordingly, it is possible to suppress the abrasion of pivot portion 117a and support groove 111b, in particular, the abrasion of support groove 111b made of a material having a low rigidity relative to cam ring 117. Therefore, it is possible to improve the endurance of the pump.

By this function, the forces acted on the inner and outer circumference sides of cam ring 117 on the pump discharge side are canceled. On the other hand, the atmospheric pressure or the suction pressure is acted through introduction passage 124 on the outer circumference region of cam ring 117 on the pump suction side which is opposite to support groove 111b. Pivot portion 117a is pressed into support groove 111b by the atmospheric pressure or the suction pressure. Accordingly, pivot portion 117a is not apart from (disengaged from) the inside surface of support groove 111b. With this, pivot portion 117a is appropriately abutted and slid on support groove 111b, and it is possible to obtain an appropriate activation of cam ring 117.

Moreover, the both of pressure chambers 131 and 132 are disposed in the region on the pump discharge side, so as to confront pump chambers 120 in the discharge region, as described above. With this, in this region, the pressure acted on the inner circumference side and the pressure acted on the outer circumference side of cam ring 117 become the discharge pressure. Accordingly, the pressure acted on the inner circumference side of cam ring 117 is substantially identical to the pressure acted on the outer circumference side of cam ring 117. Therefore, it is possible to suppress (minimize) the pressure difference between the inner circumference and the outer circumference of cam ring 117 in the discharge region. Consequently, it is possible to suppress (minimize) the leakage of the lubricating oil in the discharge region through the minute clearances between the both side surfaces of cam ring 117, bottom wall 113a of pump receiving chamber 113, and the inner side surface of cover member 112. Therefore, it is possible to cut the waste of the work of variable displacement pump 01, and to improve the efficiency of variable displacement pump 01.

As described above, in variable displacement pump 01, first and second pressure chambers 131 and 132 are disposed on the both sides of pivot portion 117a to sandwich pivot portion 117a. With this, the internal pressure of second control hydraulic chamber 132 serves to assist the urging force of coil spring 118. Accordingly, it is possible to set the urging force of coil spring 118 to a small value.

That is, by the disposition of second control hydraulic chamber 132, coil spring 118 only needs to have the urging force which ensures low pressure characteristic X, and which balances with first activation hydraulic pressure Px. Accordingly, it is possible to use the coil spring with the low load which has a spring constant smaller than that of the conventional coil spring. Consequently, it is possible to decrease the space for the disposition of coil spring 118 in pump housing 111, and to decrease the size and the weight of variable displacement pump 01. Therefore, it is possible to improve the mounting characteristic of variable displacement pump 01 on the engine.

Moreover, second pressure receiving surface 134 has the pressure receiving area smaller than the pressure receiving area of first pressure receiving surface 133. The activation hydraulic pressure of cam ring 117 is set to the two stages (steps) by second control hydraulic chamber 132. With this, it is possible to improve degree of freedom of the discharge pressure characteristic of the pump.

Furthermore, the operation of the valve timing control apparatus and the lock release hydraulic pressure of the lock mechanism are set to hydraulic pressure P1 in low pressure characteristic X. Accordingly, it is possible to improve the responsiveness of the operation of the valve timing control apparatus, like the first and second embodiments.

Moreover, first discharge pressure X is set smaller than valve opening pressure P3 of oil jet 30. Accordingly, oil jet 30 does not inject the hydraulic fluid in the engine speed which is used in the normal running of the vehicle.

Therefore, like the first and second embodiments, it is possible to suppress the discharge amount of variable displacement pump 01, to decrease the friction of various portions, and to improve the fuel consumption.

Moreover, oil jet 30 does not inject the oil of the low temperature in the cold state of the engine. Accordingly, it is possible to improve the warm-up characteristic.

Heretofore, there are proposed various pump such as the variable displacement pump for the power steering apparatus, which is arranged to control and swing the cam ring by the pressure difference between the two pressure chambers. These conventional pump is arranged to generate the pressure difference based on the pressure loss by the orifice and so on. This pressure loss decreases the pump efficiency. On the other hand, in variable displacement pump 01 according to the third embodiment, the discharge pressure is introduced into first control hydraulic chamber 131 and second control hydraulic chamber 132 without the pressure loss. Variable displacement pump 01 is arranged to generate the activation torque of cam ring 117 by the difference of the areas of the pressure receiving surfaces of first control hydraulic chamber 131 and second control hydraulic chamber 132, that is, by the difference between the areas of first pressure receiving surface 133 and second pressure receiving surface 134. Accordingly, in variable displacement pump 01, the decrease of the pump efficiency is not generated, unlike the conventional variable displacement pump. Therefore, it is possible to improve the pump efficiency relative to the conventional variable displacement pump since the pressure loss is not generated.

Moreover, variable displacement pump 01 is set to the high pressure characteristic when solenoid valve 140 is not energized. Variable displacement pump 01 has a fail-safe function to ensure necessary discharge pressure in the entire region which is used by the engine when solenoid valve 140 is in the failure state.

FIGS. 30 and 31 show variation of the third embodiment. In this variation of the third embodiment, solenoid valve 140 is a normally-closed type, unlike solenoid valve 140 of the third embodiment.

That is, solenoid valve 140 according to this variation of the third embodiment is the normally-closed type which is opposite to solenoid valve 140 of the third embodiment. As shown in FIG. 30, in the non-energization state, IN port 141a is shut off, and OUT port 141b is connected with drain portion 141c. As shown in FIG. 31, in the energization state, IN port 141a is connected with OUT port 141b. With this, variable displacement pump 01 has low pressure characteristic X when solenoid valve 140 is not energized. Variable displacement pump 01 has high pressure characteristic Y when solenoid valve 140 is energized.

By the thus-constructed configuration, when the frequency of high pressure characteristic Y is less than the frequency of low pressure characteristic X, it is possible to decrease the time duration of the energization to solenoid valve 140, and thereby to suppress the time degradation of solenoid valve 140.

Fourth Embodiment

FIGS. 32-36 show a variable displacement pump according to a fourth embodiment of the present invention. In this fourth embodiment, the dispositions of seal members 130 and 130 are varied, and solenoid valve 140 is integrally formed with the housing, unlike the third embodiment.

In this fourth embodiment, seal holding grooves 117e and 117f formed in seal constituting sections 117c and 117d of cam ring 117 are omitted, unlike the third embodiment. Alternatively, seal sliding surface 111c and 111d include seal holding grooves 111e and 111f which are identical to seal holding grooves 117e and 117f of the third embodiment, and which are formed, respectively, at positions to confront the omitted seal holding grooves 117e and 117f of the third embodiment. Seal holding grooves 111e and 111f receive, respectively, resilient members 129 and 129 and seal members 130 and 130.

Moreover, in this fourth embodiment, valve body 141 of solenoid valve 140 is integrally formed in outside surface 112b of cover member 112 to be substantially parallel with cam ring eccentric direction line N, as shown in FIGS. 32, 35 and 36. Solenoid valve 140 and the housing are integrally formed.

Solenoid valve 140 has a structure identical to that of the solenoid valve of the third embodiment. Valve element 142 is slidably received within valve body 141 integrally formed with cover member 112. An electromagnetic unit 144 is mounted in an opening portion of an one end portion which is an upper end portion of valve body 141 in FIG. 35.

By the change of these structure, on the inner side surface 112c of cover member 112, there are formed suction port 121, discharge port 122, connection groove 123 connecting discharge port 122 and bearing hole 112a, and introduction passage 125 extending from discharge port 122, like pump housing 111, as shown in FIG. 36.

This cover member 112 includes an IN port 141a which is formed at a predetermined position of introduction passage 125, and which connects the inside (pump receiving chamber 113) of pump housing 111 and the inside of valve body 141; and an OUT port 141b which is formed at a predetermined position substantially symmetrical to IN port 141a with respect to cam ring reference line M, and which serves as introduction hole 135. Valve body 111 integrally formed with cover member 112 includes a drain port 141c and a back pressure port 141d formed at predetermined positions on the circumference wall and the bottom wall of valve body 111.

Accordingly, at the swing eccentric movement of cam ring 117, seal members 130 and 130 are slidably abutted on seal surfaces 117g and 117h of cam ring 117 which is made of sintered material of iron, and which has a hardness higher than a hardness of pump housing 111 made of aluminum alloy. Therefore, it is possible to suppress the abrasion of counterpart by seal members 130 and 130. Consequently, it is possible to improve the endurance (durability) of variable displacement pump 01, relative to the third embodiment.

Moreover, in this fourth embodiment, solenoid valve 140 is integrally formed with cover member 112, that is, the housing. The entire of the hydraulic circuit of variable displacement pump 01 is formed in variable displacement pump 01. Accordingly, it is possible to decrease the size of the hydraulic pressure supply system including variable displacement pump 01 as a main device.

Fifth Embodiment

FIGS. 37-39 show a fifth embodiment of the present invention. The fifth embodiment has a basic structure identical to that of the fourth embodiment. In this fifth embodiment, a hydraulic directional control valve 150 activated by the discharge pressure is arranged to vary the discharge pressure characteristic of the pump, in place of solenoid valve 140 according to the fourth embodiment.

In this fifth embodiment, known hydraulic directional switching valve 150 of a spool type is used in place of solenoid valve 140. As shown in FIGS. 37 and 38, this hydraulic directional control valve 150 includes a valve body 151 which is formed into a substantially cylindrical shape, and which has one end opened, and the other end closed; a plug 152 which closes the opening of the one end of valve body 151; a valve element 153 which is received within valve body 151, which is arranged to be slid within valve body 151 in the axial direction, and which includes first and second land portions 153a and 153b that are located in the both end portions of valve element 153, and that define a pressure chamber 155 and a back pressure chamber 156 within valve body 151; a spring 154 which is received in back pressure chamber 156, and which is arranged to urge valve element 153 toward pressure chamber 155. When an internal pressure of pressure chamber 155 exceeds a predetermined set pressure Pz which is larger than required hydraulic pressure P1, and smaller than required hydraulic pressure P2, valve element 153 is moved toward back pressure chamber 156 against the urging force of spring 154, as shown in FIG. 38.

Valve body 151 includes an IN port 151a which is formed in the circumference wall of valve body 151 at a predetermined axial position, which penetrates the circumference wall of valve body 151, and which is connected with discharge hole 122a; an OUT port 151b which is formed in the circumference wall of valve body 151 at a predetermined axial position, which penetrates the circumference wall of valve body 151, and which is connected with introduction hole 135; and a drain port 151c which is formed in the circumference wall of valve body 151 at a predetermined axial position, which penetrates the circumference wall of valve body 151, and which is connected with suction port 121 or the outside. Moreover, valve body 151 includes a back pressure port 151d which is formed in a side wall of valve body 151 that is on the back pressure chamber 155's side, which penetrates the side wall of valve body 151, and which is connected with suction port 121 or the outside to constantly open (connect) back pressure chamber 145 to the suction pressure or the atmospheric pressure.

Plug 152 is screwed in an internal thread portion formed on an inner circumference surface of the opening of the one end portion of valve body 151. Plug 152 includes an introduction port 152a extending in the axial direction, and penetrating plug 152. The discharge pressure is constantly introduced through introduction port 152a into pressure chamber 155.

Valve element 153 includes an axially central portion which is located at a central portion of valve element 153 in the axial direction, and which has a smaller diameter; and land portions 153a and 153b which are located on both sides of the central portion, and which define an annular space 157 with valve body 151. OUT port 151b, and IN port 151a or drain port 151c are connected with each other through annular space 157.

That is, in the non-activation state of valve element 153, first land portion 153a closes IN port 151a, and OUT port 151b and drain port 151c are connected with each other through annular space 157. In the activation state of valve element 153, second land portion 153b closes drain port 151c, and IN port 151a and OUT port 151b are connected with each other through annular space 157.

Accordingly, in variable displacement pump 01 of the fifth embodiment, when the engine speed is low, IN port 151a of hydraulic directional control valve 150 is closed, so that the discharge pressure is acted only to first control hydraulic chamber 131. Consequently, when the discharge pressure reaches first activation hydraulic pressure Px as shown in FIG. 40, cam ring 117 is swung in a direction in which the eccentric amount is decreased to obtain low pressure characteristic X in which the increase of the discharge pressure becomes gradual (region T1 in FIG. 40).

Then, when the internal pressure of pressure chamber 155 reaches set pressure Pz by the increase of the discharge pressure, valve element 153 starts to be moved by the internal pressure of pressure chamber 155 in the axial direction toward back pressure chamber 156 against the urging force of spring 153. Accordingly, drain port 151c is closed by second land portion 153b, and IN port 151a is gradually opened to annular space 157. With this, IN port 151a and OUT port 151b are gradually connected with each other through annular space 157, so that the discharge pressure is gradually introduced into second control hydraulic chamber 132. Consequently, the internal pressure of second control hydraulic chamber 132 is increased, and cam ring 117 is moved in a direction in which the eccentric amount of cam ring 117 is increased. Therefore, high pressure characteristic Y to further increase the discharge pressure is attained (region T2 in FIG. 40).

In this way, by the fifth embodiment, it is possible to obtain an oil pump having a discharge pressure characteristic corresponding to the engine speed, by a lower manufacturing cost.

Moreover, the activation pressure of the valve timing control apparatus is set to the hydraulic pressure P1 in low pressure characteristic X. The valve opening pressure of oil jet 30 is set to the hydraulic pressure P3 in the hydraulic pressure characteristic Y. In this case, first activation hydraulic pressure Px is set to a hydraulic pressure sufficiently smaller than the hydraulic pressure P3. Accordingly, it is possible to decrease the consumption energy, like the first to fourth embodiments.

Moreover, in the above-described embodiments, the operation of cam ring 117 is controlled by balancing the urging force of coil spring 118 and the internal pressure of second control hydraulic chamber 132 with respect to the internal pressure of first control hydraulic chamber 131. The pressure receiving area of first pressure receiving surface 133 may be set larger than the pressure receiving area of second pressure receiving surface 134 in accordance with the specifications of the pump, and thereby coil spring 118 may be omitted. With this, the activation of cam ring 117 may be controlled only by the internal pressures (pressure difference) of pressure chambers 131 and 132.

Moreover, in the above-described embodiments, the pressure receiving area of second pressure receiving surface 134 is set smaller than the pressure receiving area of first pressure receiving surface 133. However, the pressure receiving area of second pressure receiving surface 134 may be set equal to the pressure receiving area of first pressure receiving surface 133 in accordance with the requirement of the internal combustion engine.

Moreover, the seal members are disposed to ensure the sealing ability of the control hydraulic chamber. The seal members may be omitted for the cost saving as long as the requested hydraulic pressure characteristic of the internal combustion engine is satisfied.

Moreover, the disposition of the spring receiving chamber may be varied. Set loads of the coil springs may be varied in accordance with the specifications and the size of the pump. Furthermore, the coil diameters and lengths of the coil springs may be varied.

The variable valve actuating apparatus is not limited to the valve timing control apparatus. Moreover, the present invention is applicable to, for example, a lift varying mechanism to vary a working angle (operation angle) and a lift amount of valve of the engine.

Moreover, this variable displacement pump is applicable to hydraulic equipments and so on which are other than the internal combustion engine.

In the present invention, a variable displacement pump arranged to supply a hydraulic fluid to an oil jet arranged to inject the hydraulic fluid to a piston of an internal combustion engine when a pressure of the supplied hydraulic fluid becomes equal to or greater than a predetermined pressure, the variable displacement pump includes: a pump constituting section arranged to be driven and rotated by the internal combustion engine, and thereby to discharge the hydraulic fluid entered from a suction portion to a plurality of operation chambers, from a discharge portion by volume variations of the operation chambers; a movable member arranged to decrease the flow rate of the hydraulic fluid discharged from the discharge portion by moving in one direction; a control section configured to move the movable member in the one direction by a predetermined amount when the discharge pressure of the hydraulic fluid becomes a first discharge pressure, and to further move the movable member in the one direction when the discharge pressure of the hydraulic fluid becomes a second discharge pressure larger than the first discharge pressure, the first discharge pressure being set smaller than the predetermined pressure at which the oil jet starts to inject the hydraulic fluid.

Accordingly, it is possible to suppress the energy consumption at the initial stage of the discharge of the hydraulic fluid.

In the variable displacement pump according to the present invention, the second discharge pressure is set larger than the predetermined pressure at which the oil jet starts to inject the hydraulic fluid.

Accordingly, the second discharge pressure is set larger than the predetermined pressure at which the oil jet starts to inject the hydraulic fluid. Therefore, it is possible to ensure the injection from the oil jet to the piston without being influenced by the increase of the oil temperature and the variation of the cooling state of the internal combustion engine.

In the variable displacement pump according to the present invention, the oil jet includes; a body including a hydraulic fluid supplying portion to which the hydraulic fluid is supplied, a hydraulic fluid introducing portion arranged to introduce the hydraulic fluid supplied to the hydraulic fluid supplying portion, and a valve seat formed between the hydraulic fluid supplying portion and the hydraulic fluid introducing portion; a valve element arranged to be seated on and released from the valve seat in accordance with the pressure of the hydraulic fluid supplied to the hydraulic fluid supplying portion, and thereby to open and close the hydraulic fluid supplying portion; an urging member arranged to urge the valve element in a valve closing direction in which the valve element closes the hydraulic fluid supplying portion, and to set a valve opening pressure of the valve element at which the valve element opens the hydraulic fluid supplying portion, to a value larger than the first discharge pressure; and an injection nozzle connected on a downstream side of the hydraulic pressure introducing portion, and arranged to inject the hydraulic fluid from an injection opening toward the piston.

In the variable displacement pump according to the present invention, the movable member is a cam ring having a cam surface formed on an inner circumference surface thereof; the pump forming section includes a rotor arranged to be driven and rotated by the internal combustion engine, and vanes disposed on an outer circumference portion of the rotor, and arranged to be moved in a radially inward direction or in a radially outward direction, and to be moved in the radially outward direction toward the inner circumference surface to separate the plurality of the operation chambers; and the cam ring is arranged to move to vary an eccentric amount of the cam ring with respect to a center of the rotor.

In the variable displacement pump according to the present invention, the discharged hydraulic fluid lubricates sliding portions of the internal combustion engine.

In the variable displacement pump according to the present invention, the discharged hydraulic fluid activates a valve timing control apparatus arranged to vary a relative rotational phase between a driving rotational member and a cam shaft of the internal combustion engine, and a lock mechanism of the valve timing control apparatus; and the lock mechanism has a release pressure at which a lock of the lock mechanism is released, and which is set smaller than the first discharge pressure.

The entire contents of Japanese Patent Application No. 2010-26335 filed Feb. 9, 2010 are incorporated herein by reference.

Although the invention has been described above by reference to certain embodiments of the invention, the invention is not limited to the embodiments described above. Modifications and variations of the embodiments described above will occur to those skilled in the art in light of the above teachings. The scope of the invention is defined with reference to the following claims.

Claims

1. A variable displacement pump arranged to supply a hydraulic fluid to an oil jet arranged to inject the hydraulic fluid to a piston of an internal combustion engine when a pressure of the supplied hydraulic fluid becomes equal to or greater than a predetermined pressure, the variable displacement pump comprising:

a pump constituting section arranged to be driven and rotated by the internal combustion engine, and thereby to discharge the hydraulic fluid entered from a suction portion to a plurality of operation chambers, from a discharge portion by volume variations of the operation chambers;

a movable member arranged to decrease the flow rate of the hydraulic fluid discharged from the discharge portion by moving in one direction; and

a control section configured to move the movable member in the one direction by a predetermined amount when the discharge pressure of the hydraulic fluid becomes a first discharge pressure, and to further move the movable member in the one direction when the discharge pressure of the hydraulic fluid becomes a second discharge pressure larger than the first discharge pressure,

the first discharge pressure being set smaller than the predetermined pressure at which the oil jet starts to inject the hydraulic fluid.

2. The variable displacement pump as claimed in claim 1, wherein the second discharge pressure is set larger than the predetermined pressure at which the oil jet starts to inject the hydraulic fluid.

3. The variable displacement pump as claimed in claim 1, wherein the oil jet includes;

a body including a hydraulic fluid supplying portion to which the hydraulic fluid is supplied, a hydraulic fluid introducing portion arranged to introduce the hydraulic fluid supplied to the hydraulic fluid supplying portion, and a valve seat formed between the hydraulic fluid supplying portion and the hydraulic fluid introducing portion;

a valve element arranged to be seated on and released from the valve seat in accordance with the pressure of the hydraulic fluid supplied to the hydraulic fluid supplying portion, and thereby to open and close the hydraulic fluid supplying portion;

an urging member arranged to urge the valve element in a valve closing direction in which the valve element closes the hydraulic fluid supplying portion, and to set a valve opening pressure of the valve element at which the valve element opens the hydraulic fluid supplying portion, to a value larger than the first discharge pressure; and

an injection nozzle connected on a downstream side of the hydraulic pressure introducing portion, and arranged to inject the hydraulic fluid from an injection opening toward the piston.

4. The variable displacement pump as claimed in claim 1, wherein the movable member is a cam ring having a cam surface formed on an inner circumference surface thereof; the pump forming section includes a rotor arranged to be driven and rotated by the internal combustion engine, and vanes disposed on an outer circumference portion of the rotor, and arranged to be moved in a radially inward direction or in a radially outward direction, and to be moved in the radially outward direction toward the inner circumference surface to separate the plurality of the operation chambers; and the cam ring is arranged to move to vary an eccentric amount of the cam ring with respect to a center of the rotor.

5. The variable displacement pump as claimed in claim 4, wherein the discharged hydraulic fluid lubricates sliding portions of the internal combustion engine.

6. The variable displacement pump as claimed in claim 1, wherein the discharged hydraulic fluid activates a valve timing control apparatus arranged to vary a relative rotational phase between a driving rotational member and a cam shaft of the internal combustion engine, and a lock mechanism of the valve timing control apparatus; and the lock mechanism has a release pressure at which a lock of the lock mechanism is released, and which is set smaller than the first discharge pressure.

7. A lubricating system comprising:

an oil jet arranged to inject a hydraulic fluid to a piston of an internal combustion engine when a pressure of a supplied hydraulic fluid becomes equal to or greater than a predetermined pressure; and

a variable displacement pump arranged to supply the hydraulic fluid to the oil jet, the variable displacement pump including; a pump constituting section arranged to be driven and rotated by the internal combustion engine, and thereby to discharge the hydraulic fluid entered from a suction portion to a plurality of operation chambers, from a discharge portion by volume variations of the operation chambers; a movable member arranged to decrease the flow rate of the hydraulic fluid discharged from the discharge portion by moving in one direction; and a control section configured to move the movable member in the one direction by a predetermined amount when the discharge pressure of the hydraulic fluid becomes a first discharge pressure, and to further move the movable member in the one direction when the discharge pressure of the hydraulic fluid becomes a second discharge pressure larger than the first discharge pressure,

the first discharge pressure being set smaller than the predetermined pressure at which the oil jet starts to inject the hydraulic fluid.

8. An oil jet comprising:

a body including a hydraulic fluid supplying portion to which a hydraulic fluid is supplied, a hydraulic fluid introducing portion arranged to introduce the hydraulic fluid supplied to the hydraulic fluid supplying portion, and a valve seat formed between the hydraulic fluid supplying portion and the hydraulic fluid introducing portion;

a valve element arranged to be seated on and released from the valve seat in accordance with the pressure of the hydraulic fluid supplied to the hydraulic fluid supplying portion, and thereby to open and close the hydraulic fluid supplying portion;

an urging member arranged to urge the valve element in a valve closing direction in which the valve element closes the hydraulic fluid supplying portion, and to set a valve opening pressure of the valve element at which the valve element opens the hydraulic fluid supplying portion, to a value larger than the first discharge pressure; and

an injection nozzle connected on a downstream side of the hydraulic pressure introducing portion, and arranged to inject the hydraulic fluid from an injection opening toward the piston,

the hydraulic fluid supplying portion of body receiving the supply of the hydraulic fluid from a variable displacement pump including a pump constituting section arranged to be driven and rotated by an internal combustion engine, and thereby to discharge the hydraulic fluid entered from a suction portion to a plurality of operation chambers, from a discharge portion by volume variations of the operation chambers, a movable member arranged to decrease the flow rate of the hydraulic fluid discharged from the discharge portion by moving in one direction, and a control section configured to move the movable member in the one direction by a predetermined amount when the discharge pressure of the hydraulic fluid becomes a first discharge pressure, and to further move the movable member in the one direction when the discharge pressure of the hydraulic fluid becomes a second discharge pressure larger than the first discharge pressure,

the valve element of the body being arranged to open the hydraulic fluid supplying portion to inject the hydraulic fluid to a piston of the internal combustion engine when a pressure of the supplied hydraulic fluid becomes equal to or greater than a predetermined pressure larger than the first discharge pressure.

9. The oil jet as claimed in claim 8, wherein the oil jet is arranged to start to inject the hydraulic fluid at the predetermined pressure smaller than the second discharge pressure.