Oil Pump

Abstract

An oil pump includes: a valve receiving hole including; a first port connected with the suction port, a second port connected with the switching port, a third port connected with the discharge port, and a discharge pressure introduction port arranged to receive the discharge pressure, a valve element arranged to switch a first state in which the second port and the third port are connected, and a second state in which the first port and the second port are connected; and an urging member arranged to urge the valve element toward the first end side, the pressure receiving area of the valve element and the urging force of the urging member being set so that the oil pump is switched from the first state to the second state when the pressure within the hydraulic fluid chambers to which the switching port is opened is a negative pressure.

Description

BACKGROUND OF THE INVENTION

This invention relates to an oil pump which is applied to a hydraulic source arranged to supply a hydraulic fluid to sliding portions and so on of an internal combustion engine for a vehicle.

U.S. Patent Application Publication 2008/0308062 A1 (corresponding to Japanese Patent Application Publication No. 2008-309049) discloses a conventional oil pump applied to an internal combustion engine for a vehicle. This oil pump is a variable displacement vane pump including a cam ring which is constantly urged by a spring in a direction in which the cam ring is eccentric from a center of a rotation of a rotor. An eccentric amount of the cam ring is controlled by a discharge pressure introduced into a control hydraulic chamber defined between a housing and the cam ring, so that the discharge amount becomes variable. With this, a driving torque of the pump is decreased to attain energy saving.

SUMMARY OF THE INVENTION

In recent years, it is desired to increase the discharge amount and to decrease the size of the pump by driving the conventional oil pump at a high rotational speed higher than a rotational speed of the internal combustion engine.

However, when the conventional oil pump is driven at the high rotational speed, the suction amount is not caught up, so that the cavitation is generated. Accordingly, it is not possible to ensure the sufficient discharge amount.

It is, therefore, an object of the present invention to provide an oil pump devised to solve the above-mentioned problems, to suppress a cavitation even at a high rotational speed, and to ensure a sufficient discharge amount.

According to one aspect of the present invention, an oil pump comprises: a pump constituting section arranged to continuously increase or decrease volumes of a plurality of hydraulic fluid chambers by being driven by an internal combustion engine; a port block including; a suction port opened in a region in which volumes of the hydraulic fluid chambers are increased, a switching port opened on an upstream side of a region in which the volumes of the hydraulic fluid chambers are decreased, and a discharge port opened on a downstream side of the region in which the volumes of the hydraulic fluid chambers are decreased, a valve receiving hole including; a first port formed in an inner circumference surface of the valve receiving hole, and connected with the suction port, a second port formed in the inner circumference surface of the valve receiving hole, and connected with the switching port, a third port formed in the inner circumference surface of the valve receiving hole, and connected with the discharge port, and a discharge pressure introduction port formed at a first end portion of the valve receiving hole, and arranged to receive the discharge pressure which is the hydraulic pressure within the discharge port, a valve element slidably received within the valve receiving hole, and arranged to switch a first state in which the second port and the third port are connected and a connection between the first port and the second port is restricted, and a second state in which the first port and the second port are connected and the connection between the second port and the third port is restricted; and an urging member provided on a second end side of the valve receiving hole, and arranged to urge the valve element toward the first end side of the valve receiving hole, the valve element having a pressure receiving area which is on the first end side of the valve element, and to which the discharge pressure is acted, the pressure receiving area of the valve element and the urging force of the urging member being set so that the oil pump is switched from the first state to the second state when the pressure within the hydraulic fluid chambers to which the switching port is opened is a negative pressure.

According to another aspect of the invention, an oil pump comprises: a pump constituting section arranged to continuously increase or decrease volumes of a plurality of hydraulic fluid chambers by being driven by an internal combustion engine; a port block including; a suction port opened in a region in which volumes of the hydraulic fluid chambers are increased, a switching port opened on an upstream side of a region in which the volumes of the hydraulic fluid chambers are decreased, and a discharge port opened on a downstream side of the region in which the volumes of the hydraulic fluid chambers are decreased, and a control section configured to control so that the switching port discharges the hydraulic fluid to an outside with the discharge port at a low rotational speed at which the pressure within the hydraulic fluid chamber to which the switching port is opened is a positive pressure, and to control so that the hydraulic fluid is supplied from a low pressure portion connected with the suction port, to the switching port at a high rotational speed at which the pressure within the hydraulic fluid chambers to which the switching port is opened is a negative pressure.

According to still another aspect of the invention, an oil pump comprises: a pump constituting section arranged to continuously increase or decrease volumes of a plurality of hydraulic fluid chambers by being driven by an internal combustion engine; a port block including; a suction port opened to at least a region in which volumes of the hydraulic fluid chambers are increased, and a discharge port opened to a region in which volumes of the hydraulic fluid chambers are decreased; and a control section configured to control so that, at a rotational speed at which the pressure within the hydraulic fluid chambers on an upstream side of the region in which the volumes of the hydraulic fluid chambers are decreased becomes a negative pressure, the hydraulic fluid is supplied from a low pressure portion to the hydraulic fluid chambers which are the negative pressure.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a front view showing a balancer apparatus provided integrally with an oil pump according to the present invention.

FIG. 2 is a sectional view taken along a section line A-A of FIG. 1.

FIG. 3 is a plan view showing the balancer apparatus of FIG. 1.

FIG. 4 is a bottom view showing the balancer apparatus of FIG. 1.

FIG. 5 is an exploded perspective view showing an oil pump according to a first embodiment of the present invention.

FIG. 6 is a longitudinal sectional view showing the oil pump of FIG. 5.

FIG. 7 is a sectional view showing the oil pump of FIG. 5.

FIG. 8 is a view showing a pump body of FIG. 5 as viewed from a mating surface's side with a cover member.

FIG. 9 is a view showing the cover member of FIG. 5 as viewed from a mating surface's side with the pump body.

FIG. 10 is a graph showing a relationship between spring loads of first and second springs of FIG. 7 and a swing angle of a cam ring.

FIG. 11 is a back view showing the oil pump of FIG. 5.

FIG. 12 is a sectional view taken along a section line C-C of FIG. 11.

FIG. 13 is a view showing a control valve (valve body) of FIG. 12 as viewed from a mating surface's side with the cover member.

FIGS. 14A and 14B are hydraulic circuit diagrams of the oil pump of FIG. 5. FIG. 14A shows a first state of the oil pump. FIG. 14B shows a second state of the oil pump.

FIG. 15 is a graph showing a hydraulic characteristic of the oil pump of FIG. 5.

FIGS. 16A-16E are views each showing a phase of the cam ring and a state of the hydraulic circuit in each section of FIG. 15. FIG. 16A shows a section a. FIG. 16B shows a section b. FIG. 16C shows a section c. FIG. 16D shows a section d. FIG. 16E shows a section e.

FIGS. 17A and 17B are hydraulic circuit diagrams showing an oil pump according to a second embodiment of the present invention. FIG. 17A shows a first state of the oil pump. FIG. 17B shows a second state of the oil pump.

FIGS. 18A and 18B are hydraulic circuit diagrams showing an oil pump according to a third embodiment of the present invention. FIG. 18A shows a first state of the oil pump. FIG. 18B shows a second state of the oil pump.

DETAILED DESCRIPTION OF THE INVENTION

Hereinafter, oil pumps according to first to third embodiments of the present invention are illustrated with reference to drawings. In below-described embodiments, these oil pumps are applied as an oil pump arranged to supply a lubricating oil of an engine, to sliding parts of the internal combustion engine for a vehicle, and a valve timing control apparatus configured to control an opening timing and a closing timing of an engine valve.

FIGS. 1-16 show an oil pump according to a first embodiment of the present invention. This oil pump 10 is provided integrally with a balancer apparatus 1 provided below the internal combustion engine, and arranged to decrease a second-order oscillation (secondary oscillation) of the internal combustion engine. This oil pump 10 is driven by the balancer apparatus 1.

Hereinafter, the balancer apparatus 1 is illustrated before illustration of oil pump 10. As shown in FIGS. 1-4, balancer apparatus 1 is received in an oil pan (not shown) provided below the engine. The balancer apparatus 1 includes an upper housing 2 integrally formed with the oil pan; a lower housing 3 mounted to a lower portion of the upper housing 2; a drive shaft 4 and a driven shaft 5 which are a pair of balance shafts, which are rotatably supported by the upper and lower housings 2 and 3, and which are arranged parallel with each other along an axial direction of a crank shaft 6 extending in forward and rearward directions of the engine; and a drive gear 4a and a driven gear 5a which are helical type, which are fixed, respectively, at first axial end portions (right end portions of FIG. 2) of shafts 4 and 5, and which are arranged to be rotated in opposite directions by engaging with each other. Drive shaft 4 is a drive side, and driven shaft 5 is a driven side driven by drive shaft 4. Drive shaft 4 is driven at twice the rotational speed of the crank shaft 6 by the power of crank shaft 6 which is transmitted through a chain 7. On the other hand, driven shaft 5 is driven at the same rotational speed as the drive shaft 4 in opposite direction of drive shaft 4 through gears 4a and 5a.

Oil pump 10 is mounted to a front end portion of the balancer apparatus 1. A drive shaft 14 of the oil pump 10 is connected with a second axial end portion (left end portion of FIG. 2) of driven shaft 5. With this, drive shaft 14 is driven and rotated by the power transmitted from driven shaft 5. Moreover, an oil strainer 8 is provided to lower housing 3. Oil strainer 8 is arranged to suck the hydraulic fluid stored in the oil pan. The hydraulic fluid sucked by oil strainer 8 is guided through a suction passage (not shown) provided within lower housing 3, to a suction opening 21a (described later) of oil pump 10. On the other hand, upper housing 2 includes a discharge hole 2a arranged to discharge (introduce) the hydraulic fluid to a cylinder block (not shown). The hydraulic fluid discharged from a discharge opening 22a (described later) of oil pump 10 is introduced through a discharge passage (not shown) formed within housings 2 and 3, to discharge hole 2a. Then, the hydraulic fluid is discharged through discharge hole 2a to the cylinder block.

In this way, oil pump 10 and balancer apparatus 1 are integrally formed. With this, the hydraulic fluid discharged from oil pump 10 can supplied through balancer apparatus 1 to the internal combustion engine. Accordingly, it is unnecessary to provide other (new) pipes and so on connecting oil pump 10 and the internal combustion engine. Therefore, it is possible to simplify a structure around the engine.

Next, the oil pump 10 according to the present invention are illustrated. As shown in FIGS. 5-7, the oil pump 10 includes a pump housing (corresponding to a port block); drive shaft 14; and a pump constituting section. The pump housing includes a pump body 11 that has a substantially U-shape longitudinal section, and that has a pump receiving chamber 13 in which the pump constituting section (described later) is received, and which has one end portion opened; and a cover member 12 closing the one end opening of pump body 11. Drive shaft 14 is rotatably supported by the pump housing. Drive shaft 14 penetrates a substantially center portion of pump receiving chamber 13. Drive shaft 14 is driven and rotated by balancer apparatus 1. The pump constituting section is received within pump receiving chamber 13. The pump constituting section is driven and rotated in a counterclockwise direction of FIG. 7 by drive shaft 14. With this, the pump constituting section performs a pump operation by increasing or decreasing volumes of pump chambers PR which are a plurality of operation hydraulic chambers formed within the pump constituting section.

The pump constituting section is rotatably received within pump receiving chamber 13. The pump constituting section includes a rotor 15 which is rotatably received within pump receiving chamber 13, which has a center portion connected with an outer circumference of drive shaft 14, and which has a plurality of slits 15a that are formed in an outer circumference portion of rotor 15, that extend in the radial directions; vanes 16 each received within one of the plurality of slits 15a to be moved in the radially outward direction and in the radially inward direction; a cam ring 17 which is disposed radially outside rotor 15, which is arranged to be eccentric from a center of rotor 15, and which defines pump chambers PR with the pump housing, rotor 15, and adjacent two of vanes 16 and 16; and a pair of ring members 18 and 18 which has a radius smaller than a radius of rotor 15, and which are slidably disposed, respectively, on inner circumference portions of side portions of rotor 15.

Pump body 11 is integrally made from aluminum alloy. As shown in FIGS. 5-8, pump body 11 includes an end wall 11a which is one end wall of pump receiving chamber 13; a bearing hole lib which is formed at a substantially central portion of end wall 11a to penetrate end wall 11a, and which rotatably supports a first end portion of drive shaft 14; a support groove 11c which has a substantially semi-circular transverse section, which is formed in a predetermined position of an inner circumference wall of pump receiving chamber 13, and which swingably supports cam ring 17 through a pivot pin 19. Moreover, pump body 11 includes a seal sliding surface 11d which is formed on the inner circumference surface of pump receiving chamber 13, on a negative side of a Y-axis of FIG. 7 with respect to a line (hereinafter, referred to as a cam ring reference line) connecting a center of bearing hole 11b and a center of support groove 11b. A seal member 20 (described later), disposed on an outer circumference portion of cam ring 17 are abutted and slid on the seal sliding surface lid. This seal sliding surface lid has an arc shape having a predetermined radius R1 from a center of support groove 11c. The seal sliding surface lid has a circumferential length that seal member 20 can be constantly abutted and slid on seal sliding surface lid in a range in which cam ring 17 is swung to be eccentric. When cam ring 17 is swung to be eccentric, cam ring 17 is slidingly guided along the sliding surface lid. With this, it is possible to obtain the smooth operation (eccentric swing movement) of cam ring 17.

Moreover, as shown in FIGS. 7 and 8, pump body 11 includes a suction port 21 which has a substantially arc recessed shape, which is formed on an inner side surface of end wall 11a of pump body 11 radially outside bearing hole lib, which is opened in a region (hereinafter, referred to as “volume increasing region”) that the volumes of pump chambers PR are increased by the pump operation by the pump constituting section, and into which the hydraulic fluid is introduced from balancer apparatus 1 through suction opening 21a (described later). Furthermore, pump body 11 includes a discharge port 22 which has a substantially arc recessed shape, which is formed in a region (hereinafter, referred to as “volume decreasing region”) in which the volumes of pump chambers PR are decreased, on a downstream side of the volume decreasing region, and which is arranged to introduce the hydraulic fluid discharged from pump chambers PR to discharge opening 22a (described later). Moreover, pump body 11 includes a switching port 23 which has a substantially arc recessed shape, which is formed on an upstream side of the volume decreasing region, which can switch to serve as the suction port 21 or the discharge port 22 in accordance with the hydraulic fluid pressure within the engine, that is, the discharge pressure. Discharge port 22 and switching port 23 confront suction port 21 so as to sandwich bearing hole lib.

Suction port 21 includes an induction portion 24 which is formed integrally with suction port 21, which is located in a substantially central position of suction port 21 in a circumferential direction, and which protrudes toward a first spring receiving chamber 28 (described later). Moreover, suction port 21 includes suction opening 21a which is formed near a boundary portion between induction portion 24 and suction port 21, on a start end side of suction port 21, which penetrates end wall 11a of pump body 11, and which is opened to the outside to be connected with the suction passage within balancer apparatus 1. With this, the hydraulic pressure which is sucked from the oil pan through oil strainer 8, and introduced through the suction passage to suction opening 21a is supplied through induction portion 24 and suction port 21 to the pump chambers PR positioned in the volume increasing region, by the negative pressure generated by the pump operation by the pump constituting section. Suction opening 21a and also induction portion 24 are connected with a low pressure chamber 36 (described later) formed radially outside cam ring 17. Suction opening 21a is arranged to introduce the hydraulic pressure with the low pressure which is the suction pressure, to low pressure chamber 36. In the oil pump according to the present invention, a low pressure portion is entire region which becomes the suction pressure connected with suction port 21. This low pressure portion is mainly constituted by suction port 21, induction portion 24 and low pressure chamber 36 which are adjacent to suction port 21, and so on.

Discharge port 22 includes a discharge opening 22a which is located on a terminal end side of discharge port 22, which penetrates end wall 11a of pump body 11, and which is opened to the outside to be connected with the discharge passage within the balancer apparatus 1. With this, the hydraulic fluid pressurized by the pump operation by the pump constituting section, and discharged to discharge port 22 is supplied from discharge opening 22a through the discharge passage, and an oil main gallery (not shown) provided in the cylinder block, to the sliding portions within the engine, the valve timing control apparatus, and so on. In this case, discharge opening 22a is provided to protrude in the radially outward direction with respect to the circumferential ends of discharge port 22. An outside half of discharge opening 22 is connected through a first connection passage 38 (described later) formed within cam ring 17, to a discharge port 32 (a second connection hole 32a) (described later) formed in cover member 12.

Near a start end portion of discharge port 22, there is formed a connection groove 25 which connects discharge port 22 and bearing hole 11b. The hydraulic fluid is supplied through this connection groove 25 to bearing hole 11b, and side portions of rotor 15 and vanes 16. With this, it is possible to ensure good lubricating characteristics of the sliding portions. This connection groove 25 is formed so that connection groove 25 does not correspond to the movement directions of vanes 16 in the radially inward direction and in the radially outward direction. With this, when vanes 16 are moved in the radially inward direction and in the radially outward direction, it is possible to prevent vanes 16 from dropping into connection groove 25. Moreover, at the start end portion of discharge port 22, there is formed one end of an inside passage 26 which is opened to the start end portion of discharge port 22. Inside passage 26 is formed within end wall 11a of pump body 11. Inside passage 26 connects discharge port 22 and control hydraulic chambers 37 (described later). A part of the hydraulic fluid discharged to discharge port 22 is introduced through inside passage 26 to control hydraulic chamber 37.

Switching port 23 includes a connection portion 27 integrally formed at a terminal end portion of switching port 23 to protrude in the radially outward direction. Switching port 23 is connected through a second connection passage 39 (described later) formed within cam ring 17, to switching port 33 (a third connection hole 33a) of cover member 12.

As shown in FIGS. 5 and 9, cover member 12 is formed into a substantially plate shape. Cover member 12 is mounted to the opening end surface of pump body 11 by a plurality of bolts B1. Cover member 12 includes a bearing hole 12a which is formed in a position to confront bearing hole 11b of pump body 11, which rotatably supports a second end portion of drive shaft 14, and which penetrates cover member 12. On an inner side surface of cover member 12 which constitutes the other end wall of pump receiving chamber 13, there are formed a suction port 31, a discharge port 32, and a switching port 33 which are located in positions to confront ports 21-23 of pump body 11, and which have, respectively, structures substantially identical to ports 21-23. Suction port 31 includes a first connection hole 31a which is formed at a terminal end portion of suction port 31, and which is connected with a first port 46 (described later) of a control valve 40. Discharge port 32 includes a second connection hole 32a formed at a substantially central portion of discharge port 32, and connected with a second port 47 (described later) of control valve 40. Switching port 33 includes a third connection hole 33a formed at a terminal end portion of switching port 33, and connected with a third port 48 (described later) of control valve 40. First to third connection ports 31a-33a penetrates cover member 12.

As shown in FIG. 2, drive shaft 14 includes a first axial end portion 14a penetrating end wall 11a of pump body 11, and confronting the outside. First axial end portion 14a of drive shaft 14 is connected with the second axial end portion of driven shaft 5 of balancer apparatus 1 so as to rotate as a unit with driven shaft 5. Drive shaft 14 rotates rotor 15 in the counterclockwise direction of FIG. 6 by a rotational force transmitted from driven shaft 5. In this case, a line (hereinafter, referred to as “cam ring eccentric direction line”) N which passes a center of drive shaft 14, and which is perpendicular to line M is a boundary between the volume increasing region and the volume decreasing region. When control valve 40 is in the first state (described later) which is a normal state, an X-axis negative side of line N is a suction region, and an X-axis positive side of line N is a discharge region.

As shown in FIGS. 5-7, rotor 15 includes the plurality of slits 15a each of which extends from the center in the radially outside direction. Each of slits 15a includes a back pressure chamber 15b which has a substantially circular cross section, which is formed at an inside base end portion of the each of slits 15a, and into which the discharge hydraulic pressure is introduced. Each of vanes 16 is arranged to be pushed in the outside direction by the centrifugal force by the rotation of rotor 15 and the pressure within the corresponding back pressure chamber 15b.

When rotor 15 is rotated, tip end surfaces (radially outer surfaces) of vanes 16 are abutted and slid on an inner circumference surface of cam ring 17, and base end surfaces (radially inner surfaces) of vanes 16 are abutted and slid on outer circumference surfaces of ring members 18 and 18. That is, vanes 16 are arranged to be moved in the radially outward direction of rotor 15 by ring members 18 and 18. Even when the engine speed is low and the centrifugal force and the pressure of back pressure chambers 15b are small, the tip ends of vanes 16 are abutted and slid on the inner circumference surface of cam ring 17, so that pump chambers PR are liquid-tightly separated.

Cam ring 17 is made from a sintered metal, and integrally formed into a substantially cylindrical shape. Cam ring 17 includes a pivot portion 17a which is formed in a predetermined position on an outer circumference portion of cam ring 17, which is a substantially arc recessed groove, which is formed by cutting in the axial direction, and which constitutes an eccentric swing point by mounting on a pivot pin 19. Moreover, cam ring 17 includes an arm portion 17b which is located on an opposite side of the center of cam ring 17 with respect to pivot portion 17a, which links with springs 34 and 35 (described later), and which protrudes in the radial direction. On both side portions of arm portion 17b in the pivot direction, there are disposed first spring 34 which is set to a predetermined spring constant, and a second spring 35 which is set to a spring constant smaller than the spring constant of first spring 34. First spring 34 and second spring 35 confront each other to sandwich arm portion 17b of cam ring 17. Arm portion 17b includes a press protruding portion 17c which is formed into a substantially arc raised shape, and which is formed on one side portion of arm portion 17b in the pivot direction; and a press protrusion 17d which extends toward second spring receiving chamber 29, which is formed on the other side portion of arm portion 17b in the pivot direction, and which has a length longer than a thickness (width) of a restriction portion 30 (described later). Press protruding portion 17c is constantly abutted on a tip end portion of first spring 34. Press protrusion 17d is constantly abutted on a tip end portion of second spring 35. With this, arm portion 17b and springs 34 and 35 are linked with each other.

As shown in FIGS. 7 and 8, pump body 11 includes first and second spring receiving chambers 28 and 29 which are formed within pump body 11 in positions to confront support groove 11c, which receive first and second springs 34 and 35 respectively, which extend in a Y-axis direction of FIG. 7, and which are adjacent to pump receiving chamber 13. First spring 34 is mounted within first spring receiving chamber 28 between an end surface of first spring receiving chamber 28 and arm portion 17b (press protruding portion 17c) so that first spring 34 has a predetermined set load W1. Second spring 35 is mounted within second spring receiving chamber 29 between an end surface of second spring receiving chamber 29 and arm portion 17b (press protrusion 17d) so that second spring 35 has a predetermined set load W2. Second spring 35 has a wire diameter smaller than that of first spring 34. Pump body 11 includes a restriction portion 30 which is a stepped shape to decrease a diameter thereof, and which is positioned between first and second spring receiving chambers 28 and 29. The other side portion (on the lower side of FIG. 7) of arm portion 17b is arranged to be abutted on one side portion (on the upper side of FIG. 7) of restriction portion 30 so as to restrict a pivot movement region of arm portion 17b in the counterclockwise direction. On the other hand, the end of second spring 35 is arranged to be abutted on the other side portion (on the lower side of FIG. 7) of restriction portion 30 so as to restrict a maximum extension amount of second spring 35.

In this way, cam ring 17 is constantly urged through arm portion 17b in a direction to increase the eccentric amount of cam ring 17 (in the counterclockwise direction of FIG. 7) by a resultant force W0 of set loads W1 and W2 of springs 34 and 35, that is, the urging force of first spring 34 which exerts a relatively larger spring load. With this, as shown in FIG. 7, press protrusion 17d of arm portion 17b enters second spring receiving chamber 29 in the inactive state, and compresses second spring 35. The other side portion of arm portion 17b is pressed on the one side portion of restriction portion 30. With this, cam ring 17 is restricted in a position in which the eccentric amount of cam ring 17 is maximized.

As shown in FIG. 7, cam ring 17 includes a seal forming portion 17e which is formed on an outer circumference portion of cam ring 17, which protrudes in the outward direction, and which has a substantially rectangular cross section. Seal forming portion 17e includes a seal surface 17f that has a substantially arc shape concentric with seal sliding surface 11d, and that confronts seal sliding surface 11d of pump body 11; and a seal holding groove 17g which is formed on seal surface 17f of seal forming portion 17e, which has a substantially rectangular cross section, and which is formed by cutting to extend in the axial direction. A seal member 20 is received and held within this seal holding groove 17g. Seal member 20 is slid on seal sliding surface 11d at the eccentric swing movement of cam ring 17. Seal surface 17f has a predetermined radius R2 slightly smaller than radius R1 of seal sliding surface 11d. With this, there is formed a predetermined minute clearance.

Seal member 20 is made, for example, from a fluorine resin having a low frictional characteristic. Seal member 20 has an elongated linear shape extending in the axial direction of cam ring 17. A resilient member 20a made from a rubber is disposed on a bottom portion of seal holding groove 17g. Seal member 20 is pressed on seal sliding surface lid by a resilient force of resilient member 20a to liquid-tightly separate between seal sliding surface 11d and seal surface 17f.

Moreover, spring receiving chambers 28 and 29 are connected through introduction portion 24 and suction port 21 to low pressure chamber 36. Cam ring 17 is constantly pressed on pivot pin 19 by the hydraulic fluid pressure within low pressure chamber 36. Moreover, there is formed control hydraulic chamber 37 which is located radially outside cam ring 17 and switching port 23, which is separated by pivot pin 19 and seal member 20, and which serves for the eccentric swing movement control of cam ring 17.

The hydraulic fluid within discharge port 22 is constantly introduced through inside passage 26 to this control hydraulic chamber 37. Seal forming portion 17e of cam ring 17 includes a pressure receiving surface 17h which confronts control hydraulic chamber 37, and which is formed on a side surface of seal forming portion 17e. Control hydraulic chamber 37 is arranged to act the pump discharge pressure on this pressure receiving surface 17h of seal forming portion 17e, and thereby to provide, to cam ring 17, a swing movement force (movement force) in a direction in which the eccentric amount of cam ring 17 is decreased (in the clockwise direction of FIG. 7). That is, this control hydraulic chamber 37 constantly urges cam ring 17 through pressure receiving surface 17h in a direction in which the center of cam ring 17 approaches the center of rotor 15 so that cam ring 17 becomes concentric with rotor 15. With this, control hydraulic chamber 37 performs a movement amount control of cam ring 17 in a concentric direction with respect to rotor 15.

In the thus-constructed oil pump 10, the urging force in the eccentric direction by the spring load of first spring 34, the urging force in the concentric direction by the spring force of second spring 35 and the internal pressure of control hydraulic chamber 37 are balanced in a predetermined force relationship. When the urging force by the internal pressure of control chamber 37 is smaller than the resultant force (total force) W0 (=W1-W2) of first and second springs 34 and 35 which is a difference between set load W1 of first spring 34 and set load W2 of second spring 35, cam ring 17 is brought to the maximum eccentric state as shown in FIG. 7. When the urging force by the internal pressure of control chamber 37 becomes greater than resultant force W0 of the set loads of springs 34 and 35 in accordance with the increase of the discharge pressure, cam ring 17 is moved in the concentric direction by this discharge pressure.

Hereinafter, the relationship between spring load W of springs 34 and 35 and the swing angle (movement amount) X of cam ring 17 is illustrated. As shown in FIG. 10, when the urging force by the internal pressure of control hydraulic chamber 37 becomes equal to (reaches) resultant force W0 of set loads W1 and W2 of springs 34 and 35 which corresponds to an urging force by a first operation hydraulic pressure Pf (described later), in a position X1 in which cam ring 17 becomes the maximum eccentric state, first spring 34 is compressed, and second spring 35 starts to expand, so that cam ring 17 is moved in the concentric direction. Then, when second spring 35 is abutted on restriction portion 30 by the increase of the urging force by the internal pressure of control hydraulic chamber 37 by the increase of the discharge pressure of the pump, the assist function by second spring 35 disappears. With this, the movement of cam ring 17 in the concentric direction is stopped (position X2 in FIG. 10). Then, when the urging force by the internal pressure of control hydraulic chamber 37 becomes equal to spring load Wx of first spring 34 which corresponds to the urging force by a second operation hydraulic pressure Ps (described later) by the further increase of the discharge pressure, first spring 34 starts to be further compressed, and cam ring 17 is further moved in the concentric direction (a position X3 in FIG. 10).

Moreover, as shown in FIGS. 5 and 11-13, oil pump 10 includes a control valve (control section) 40 which is provided on a back portion of cover member 12, which is controlled by the discharge pressure introduced through first connection passage 38 formed within cam ring 17, and which is arranged to control a switching of a function of switching port 23 in accordance with an axial position of a valve element 43.

As shown in FIG. 12, control valve 40 includes a valve body 41 which is formed into a substantially cylindrical shape, and which has a first end portion formed to be opened, and a second end portion formed to be closed; a plug 42 closing the opened first end portion of valve body 41; a valve element 43 which is received in an inner circumference of valve body 41 to be slid in the axial direction, which includes a first land portion 43a and a second land portion 43b that are formed at both end portions of valve element 43, and that are abutted on and slid on the inner circumference surface of valve body 41; and a valve spring 44 which has a predetermined set load Wk identical to a port switching hydraulic pressure Pk (described later), which is mounted between plug 42 and valve element 43 on the inner circumference of valve body 41 on the first end portion's side of valve body 41.

Valve body 41 includes a valve element receiving portion 41a which has a diameter substantially identical to diameters of land portions 43a and 43b of valve element 43, and which receives valve element 43; and a back pressure chamber forming portion 41b which has a stepped shape that the diameter is decreased with respect to valve element receiving portion 41a, and which forms a back pressure chamber 45 therein by being separated by second land portion 43b of valve element 43. Valve body 41 is fixed on a back surface of cover member 12 by a plurality of bolts B2. Moreover, valve body 41 includes a first port 46, a second port 47, and a third port 48 which are formed in a circumferential wall of valve element receiving portion 41a to penetrate valve body 41. First port 46 is connected to first connection hole 31a through a suction pressure introduction groove 46a formed on an inner side surface of valve body 41. Second port 47 is connected to second connection hole 32a through a discharge pressure introduction groove 47a formed on the inner side surface of valve body 41. Third port 48 is opened directly to third connection hole 33a to be connected to third connection hole 33a. Furthermore, valve body 41 includes a back pressure port (discharge pressure introduction port) 49 which is formed in the circumferential wall of valve element receiving portion 41a to penetrate valve body 41, which is constantly connected to second port 47 through a connection groove 49a formed on the inner side surface of valve body 41, and into which the hydraulic fluid (the discharge pressure) discharged to discharge ports 22 and 32 is constantly introduced through connection hole 32a.

Valve element 43 includes an axial intermediate portion which has a diameter smaller than diameters of both end portions. There is separated an annular space 50 formed between this axial intermediate portion of valve element 43 and valve body 41 by both land portions 43a and 43b. Third port 48 is connected with first port 46 or second port 47 through annular space 50.

In the thus-constructed control valve 40, when the discharge pressure introduced into back pressure chamber 45 is low and the urging force by the internal pressure of this back pressure chamber 45 is smaller than set load Wk of valve spring 44 as shown in FIG. 14A, valve element 43 (second land portion 43b) is pressed on stepped portion 41c of valve body 41 by the urging force of valve spring 44. With this, first port 46 is closed by first land portion 43a, and second port 47 is connected with third port 48 through annular space 50. Accordingly, switching ports 23 and 33 function as the discharge ports (corresponding to a first state).

On the other hand, when the discharge pressure introduced into back pressure chamber 45 is increased by the increase of the engine speed, that is, by the increase of the rotational speed of oil pump 10 and the urging force by the internal pressure of back pressure chamber 45 becomes greater than set load Wk of valve spring 44 as shown in FIG. 14B, valve element 43 is moved toward the first end's side of valve body 41 (the plug 42's side) by the urging force by the discharge pressure against the urging force of valve spring 44. With this, first port 46 is connected through annular space 50 to third port 48, and second port 47 is connected with back pressure port 49. Accordingly, switching ports 23 and 33 function as the suction ports (corresponding to second state).

In this case, in the port switching control of control valve 40, the pressure receiving area of valve element 43 (second land portion 43b) and set load Wk of valve spring 44 are set so as to balance the urging force by valve spring 44, and the urging force by back pressure chamber 45 based on the discharge pressure immediately before the internal pressures of switching ports 23 and 33 become the negative pressure. In a case in which the internal pressures of switching ports 23 and 33 becomes the negative pressure in the first state, the oil pump is already switched from the first state to the second state.

Next, an operation of oil pump 10 according to the embodiment of the present invention is illustrated with reference to FIGS. 15 and 16.

First, a necessary hydraulic pressure of the internal combustion engine which is reference (basis) for the discharge pressure control of oil pump 10 is illustrated with reference to FIG. 15. A reference symbol P1 in FIG. 15 is a first engine request hydraulic pressure which corresponds to a request pressure of a valve timing control apparatus arranged to improve the fuel consumption when the valve timing control apparatus is employed. A reference symbol P2 is a second engine request hydraulic pressure which is a request pressure of an oil jet arranged to cool the piston when the oil jet is employed. A reference symbol P3 is a third engine request pressure which is needed for lubricating the bearing portions of the crank shaft at the high engine speed. A curve line E (two-dot chain line) of FIG. 15 connecting these points P1-P3 represents an ideal necessary hydraulic pressure (discharge pressure) according to engine speed R of the internal combustion engine. Moreover, a reference symbol Pf in FIG. 15 represents first operation hydraulic pressure by which cam ring 17 starts to be swung by the urging force by the internal pressure of control hydraulic chamber 37 against resultant force W0 of springs 34 and 35. A reference symbol Ps represents second operation hydraulic pressure by which cam ring 17 starts to be further swung by the urging force by the internal pressure of control hydraulic chamber 37 against spring load W1 of first spring 34. A reference symbol Pk set between the operation hydraulic pressures Pf and Ps represents port switching hydraulic pressure for switching a connection of switching ports 23 and 33.

That is, in a section a of FIG. 15 which corresponds to the idling state of the engine, oil pump 10 is held in a state in which the eccentric amount of cam ring 17 is maximum as shown in FIG. 16A, by the urging force by resultant force W0 of springs 34 and 35, that is, by the relatively larger urging force by the spring force of first spring 34. With this, in this section a, the discharge pressure (the hydraulic pressure within the engine) P is sufficiently smaller than port switching hydraulic pressure Pk. Valve element 43 of control valve 40 is positioned nearest to back pressure chamber 45 (valve element 43 is positioned in the rightmost position of FIG. 16A), and switching ports 23 and 33 are connected, respectively, with discharge ports 22 and 32. With this, discharge pressure P has a characteristic to increase to be substantially proportional to engine speed R.

Then, when engine speed R increases and discharge pressure P becomes equal to first operation hydraulic pressure Pf set to a value higher than first engine request hydraulic pressure P1, cam ring 17 starts to be moved in the concentric direction against the urging force of first spring 34 by the urging force by the internal pressure of control hydraulic chamber 37 and the urging force of second spring 35, as shown in FIG. 16B. With this, the eccentric amount of cam ring 17 is gradually decreased, so that the increase of the discharge amount is limited. Accordingly, the increase of the discharge pressure P by the increase of engine speed R is suppressed (the section b in FIG. 15). Then, when second spring 35 is expanded by the movement of cam ring 17 and the tip end of second spring 35 is abutted on restriction portion 30 as shown in FIG. 16C, the urging assist function by second spring 35 disappears, so that the movement of cam ring 17 is stopped. Accordingly, discharge pressure P is again increased by the increase of engine speed R to be substantially proportional to engine speed R (the section c in FIG. 15).

Oil pump 10 according to the embodiment of the present invention is arranged to be driven by the rotational speed of balancer apparatus 1 (driven shaft 5), that is, twice the rotational speed of the conventional rotational speed, in place of the rotational speed of the internal combustion engine (the crank shaft) in the conventional oil pump. In this oil pump 10, the cavitation may be generated in the region of engine speed R greater than a predetermined engine speed Rk. With this, it is not possible to increase discharge pressure P to be proportional to the increase of engine speed R. That is, the rotational speed of rotor 15 is too fast due to twice the rotational speed of the conventional rotational speed. Accordingly, a time period during which pump chambers PR are positioned in the region (volume increasing region) of suction ports 21 and 31 becomes too short. With this, pump chambers PR are moved to the region (the upstream side of the volume decreasing region) of switching ports 23 and 33 connected with discharge ports 22 and 32, in a state in which pump chambers PR does not suck the sufficient hydraulic fluid. When pump chambers PR are moved to the region of switching ports 23 and 33, pump chambers PR confronting switching ports 23 and 33 are in the negative pressure state. Accordingly, these pump chambers PR cannot discharge the hydraulic fluid to switching ports 23 and 33. Consequently, it is not possible to increase the discharge amount in accordance with the increase of engine speed R.

The above-described problem generates prominently in a vane type pump such as oil pump 10. For example, in case of a trochoid type pump, the discharge amount is increased in accordance with the increase of the engine speed even after the generation of the cavitation. Accordingly, the above-described problem that the discharge amount is deficient hardly generate.

In oil pump 10 according to the embodiment of the present invention, when discharge pressure P becomes, by the increase of engine speed R, equal to port switching hydraulic pressure Pk corresponding to the discharge pressure at engine speed Rk at which the cavitation may be generated, the urging force by the internal pressure of back pressure chamber 45 becomes greater than set load Wk of valve spring 44, so that valve element 43 is moved to the plug 42's side against set load Wk of valve spring 44, as shown in FIG. 16D. With this, control valve 40 is switched from the first state to the second state. Consequently, first port 46 is connected through annular space 50 to third port 48. Therefore, the hydraulic fluid is introduced from suction ports 21 and 31 through third connection hole 33a and second connection passage 39 to switching ports 23 and 33. Switching ports 23 and 33 function as the suction ports, in addition to suction ports 21 and 31.

In this way, switching ports 23 and 33 function as the suction ports. With this, the suction section (suction region) is enlarged by the suction ports 21 and 31, and also switching ports 23 and 33. Accordingly, the time period during which pump chambers PR are positioned in the suction region is lengthened. Consequently, it is possible to supply the sufficient hydraulic fluid to the pump chambers PR which become the negative pressure on the upstream side (the regions of switching ports 23 and 33) of the volume decreasing region since the hydraulic fluid is not sufficiently supplied only by the regions of suction ports 21 and 31 as mentioned above. Therefore, it is possible to suppress the generation of the cavitation.

While control valve 40 is switched from the first state to the second state, the connection amount between first port 46 and second port 47 within control valve 40 is gradually decreased. Accordingly, the inherent discharge amount (the discharge amount per one rotation of the pump) is decreased. Therefore, the increase amount of discharge pressure P temporarily remains substantially the same level even when engine speed R is increased. After the port switching is finished, discharge pressure P is increased to be substantially proportional to engine speed R in accordance with the increase of engine speed R.

In this case, the discharge region is decreased by the switching of control valve 40 to the second state. Pump chambers PR located in the discharge region discharge the hydraulic fluid to discharge ports 22 and 32 in a state in which the volumes of these pump chambers PR are decreased. With this, the theoretical discharge amount is decreased, relative to a case in which the control valve 40 is in the first state. However, the actual discharge amount is increased since the cavitation is not generated.

Next, when engine speed R further increases and discharge pressure P becomes equal to second operation hydraulic pressure Ps set greater than third engine request hydraulic pressure P3, cam ring 17 starts to be further moved in the concentric direction against spring load Wx of first spring 34 by the urging force by the internal pressure of control hydraulic chamber 37, as shown in FIG. 16E. The increase of the discharge amount is limited as the eccentric amount of cam ring 17 is decreased. With this, the increase of discharge pressure P based on the increase of engine speed R is suppressed (the section e in FIG. 15).

In this way, in oil pump 10 according to the embodiment of the present invention, the swing movement of cam ring 17 is controlled so that discharge pressure P is increased in the multistep manner by springs 34 and 35 and control valve 40. With this, it is possible to obtain a characteristic corresponding to a request hydraulic pressure (curve line E) of the internal combustion engine, without increasing discharge pressure P uselessly.

Moreover, in oil pump 10 according to the embodiment of the present invention, switching ports 23 and 33 are not merely connected with suction ports 21 and 31 at the high rotational speed that the discharge pressure P is greater than port switching hydraulic pressure Pk. Oil pump 10 is arranged to suck more hydraulic fluid at the high rotational speed by suppressing the cavitation. With this, oil pump 10 ensures the appropriate discharge amount. Oil pump 10 hardly performs the waste work, and does not suffer from the deterioration of the fuel consumption.

In this way, switching ports 23 and 33 can supply the hydraulic fluid by switching control valve 40 at the high rotational speed at which pump chambers PR located on the upstream side (the region of switching ports 23 and 33) of the volume decreasing region become the negative pressure. With this, the suction amount is increased. Accordingly, it is possible to suppress the cavitation at the high rotational speed, and to ensure the sufficient discharge amount.

That is, it is possible to suck the hydraulic fluid from switching ports 23 and 33 at the high rotational speed. With this, the region (suction region) in which the hydraulic fluid is sucked is enlarged, it is possible to ensure the longer time period (suction time) during which the pump chambers PR are positioned in the suction region, to ensure the sufficient suction amount, and to suppress the cavitation.

Moreover, in this oil pump 10, when cam ring 17 is moved in the concentric direction, the side surfaces of arm portion 17b cover parts of suction ports 21 and 31. With this, the resistance is generated at the suction. Accordingly, it is possible to effectively suppress the cavitation.

Moreover, oil pump 10 is rotated at twice the speed of the conventional rotational speed (the rotational speed of the crank shaft). With this, it is possible to have the discharge ability identical to the conventional pump apparatus by half of the pump volumes of the conventional pump apparatus. Accordingly, it is possible to considerably decrease the size of the oil pump.

Moreover, in this embodiment, the existing balancer apparatus 1 provided to the internal combustion engine is used for obtaining twice the speed of the conventional rotational speed. Accordingly, it is possible to suppress the increase of the cost without providing a new driving source.

FIGS. 17A and B are views showing an oil pump according to a second embodiment of the present invention. In the oil pump according to the second embodiment, there is provided a normally-closed solenoid valve SV disposed between back pressure port 49 of control valve 40 and discharge ports 22 and 32, and arranged to be actuated based on an excitation current from an ECU 60 mounted on the vehicle in accordance with a condition of the vehicle. Solenoid valve SV electrically performs the port switching control. FIG. 17A shows the first state in which the excitation current is not applied to solenoid valve SV. FIG. 17B shows the second state in which the excitation current is applied to solenoid valve SV.

That is, solenoid valve SV is basically controlled by engine speed R of the internal combustion engine which is sensed by a predetermined sensor and so on. When engine speed R is smaller than predetermined rotational speed Rk (cf. FIG. 15), the excitation current is not applied from ECU 60. The hydraulic fluid introduced into back pressure port 49 is drained. The discharge pressure P is not introduced into back pressure chamber 45. Consequently, valve element 43 is pressed on stepped portion 41c of valve body 41 by the urging force of valve spring 44. Control valve 40 is held in the first state (cf. FIG. 17A).

On the other hand, when engine speed R of the internal combustion engine becomes equal to or greater than predetermined rotational speed Rk, the excitation current is applied from ECU 60, so that solenoid valve SV is opened. Discharge pressure P equal to or greater than port switching hydraulic pressure Pk is introduced into back pressure chamber 45. Consequently, valve element 43 is moved to the first end side (the plug 42's side) by the introduced pressure against the urging force of valve spring 44. Control valve 40 is switched to the second state (cf. FIG. 17B)

In this way, in the oil pump according to the second embodiment of the present invention, the port switching control by the control valve 40 is performed by using solenoid valve SV in the electrical manner. Accordingly, this oil pump according to this embodiment does not suffer from influence and so on of the variation of the hydraulic pressure due to the abrasion of various parts of pump 10 and change of the kind of the hydraulic fluid. Consequently, it is possible to constantly appropriately perform the port switching control, and to more surely suppress the generation of the cavitation.

In the oil pump according to the embodiment, solenoid valve SV is disposed between back pressure port 49 of control valve 40 and discharge ports 22 and 23 to control the pressure introduced into back pressure chamber 45 of control valve 40. Moreover, solenoid valve SV may directly drive valve element 43 of control valve 40, in addition to the above-described structure. In this case, it is also possible to attain the above-described effects.

FIGS. 18A and 18B are an oil pump according to a third embodiment of the present invention. In the oil pump according to the third embodiment, solenoid valve SV in the oil pump according to the second embodiment is changed to a linear solenoid valve RSV. Moreover, an orifice 61 is provided between linear solenoid valve RSV and discharge ports 22 and 32. Linear solenoid valve RSV controls the drain amount to control the introduction pressure acted to back pressure chamber 45. FIG. 18A shows the first state in which the drain amount of linear solenoid valve RSV is maximized. FIG. 18B shows the second state in which the drain amount of linear solenoid valve RSV becomes zero.

Linear solenoid valve RSV includes a valve body 51 which is formed into a substantially cylindrical shape that has a first end portion formed to be opened, and a second end portion formed to be closed; a valve element 52 which is received on the inner circumference of valve body 51 to be slid in the axial direction, and which includes first and second land portions 52a and 52b abutted and slid on the inner circumference surface of valve body 51; a valve spring 53 which is received within a back pressure chamber 55 separated by second land portion 52b of valve element 52 on the second end side of valve body 51, and which is arranged to urge valve element 52 to the first end side of valve body 51; and an electromagnetic unit 54 which is mounted on the open first end portion of valve body 51, and which is arranged to move valve element 52 to the second end side of valve body 51 in the axial direction against the urging force of valve spring 53 by moving a rod 54b by energization.

Valve body 51 includes an IN port 51a, an OUT port 51b, and a drain port 51c which are formed on a circumferential wall of valve body 51 to penetrate the circumferential wall of valve body 51. IN port 51a is connected with switching ports 23 and 33. OUT port 51b is connected with back pressure port 49. Drain port 51c is connected with suction port 21 and 31 or the outside, and arranged to discharge the hydraulic fluid within an annular space 56 (described later).

Valve element 52 includes an axial intermediate portion which has a diameter smaller than diameters of the both end portions. Valve body 51 includes an annular space 56 separated by both land portions 52a and 52b and the axial intermediate portion of valve element 52, and which is located between valve body 51 and valve element 52. Ports 51a-51c are connected with each other through annular space 56. Second land portion 52b of valve element 52 opens and closes drain port 51c, and accordingly the hydraulic fluid pressure introduced from annular space 56 through OUT port 51b to back pressure port 49 is controlled.

Electromagnetic unit 54 includes a coil unit 54a having a bobbin wound by a coil, a york mounted on the bobbin; an armature (not shown) which is made from a magnetic material, which is disposed radially inside coil unit 54a, and which is arranged to be moved in the axial direction; and a rod 54b which is connected with the armature, and which is arranged to be moved in forward and rearward directions in accordance with excitation state of the coil.

In the thus-constructed oil pump, linear solenoid valve RSV is applied with the maximum excitation current from ECU 60 when engine speed R of the internal combustion engine is lower than predetermined engine speed Rk. Rod 54b is moved out by the maximum distance, as shown in FIG. 18A. The discharge amount of drain port 51c is maximized. Consequently, the hydraulic pressure enough to move valve element 43 against the urging force of valve spring 44 is not introduced into back pressure chamber 45 of control valve 40. Valve element 43 is pressed on stepped portion 41c of valve body 41 by the urging force. Control valve 40 is held in the first state (cf. FIG. 18A).

On the other hand, when engine speed R of the internal combustion engine increases equal to or greater than predetermined engine speed Rf corresponding to first operation hydraulic pressure Pf, the excitation current from ECU 60 is gradually decreased in accordance with the increase of engine speed, so that the open area of drain port 51c is gradually decreased. With this, the internal pressure of back pressure chamber 45 is gradually increased. When the internal pressure of back pressure chamber 45 becomes equal to port switching hydraulic pressure Pk, valve element 43 starts to move toward the first end side (the plug 42's side). The port switching control is started. Then, when the excitation current from ECU 60 is minimized and drain port 51c is fully closed, control valve 40 is completely shifted from the first state to the second state (cf. FIG. 18B).

As mentioned above, the oil pump according to this embodiment employs linear solenoid valve RSV arranged to perform the introduction pressure control of control valve 40 to back pressure chamber 45. With this, it is possible to attain the same effect as the oil pump according to the second embodiment. Moreover, the linear solenoid valve RSV can perform the port switching gradually, unlike the solenoid valve SV arranged to perform the ON-OFF control only. Accordingly, it is possible to suppress the variation of the discharge pressure based on the variation of the discharge amount at the port switching.

In the oil pump according to the embodiment of the present invention, linear solenoid valve RSV is disposed between back pressure port 49 of control valve 40 and discharge ports 22 and 32 so as to control the pressure introduced into back pressure chamber 45 of control valve 40. In addition to the above-described structure, linear solenoid valve RSV may directly drive the valve element 43 of control valve 40. In this case, it is also possible to attain the above-described same effects as the second embodiment.

The present invention is not limited to the structures according to the embodiments. For example, engine request hydraulic pressures P1-P3, first and second operation hydraulic pressures Pf and Ps, and port switching hydraulic pressure Pk may be freely varied in accordance with specifications of the internal combustion engine, the valve timing control apparatus, and so on of a vehicle on which oil pump 10 is mounted.

Moreover, it is not necessary to use balancer apparatus 1 as the driving source of oil pump 10. It is possible to use other apparatuses as long as the other apparatuses can drive at a rotational speed higher than the rotational speed of the internal combustion engine. Moreover, this other apparatuses may be existing apparatus which is mounted on the vehicle body, and may be a new apparatus.

Moreover, in the embodiments, the suction region is enlarged by connecting switching ports 23 and 33 through control valve 40 to suction ports 21 and 31. Moreover, the suction region may be continuously enlarged so as to ensure the pump suction amount at the high rotational speed, without uniformly enlarging the suction region by combining the divided ports.

That is, discharge ports 22 and 32 and switching ports 23 and 33 are formed, respectively, as one continuous discharge port. One (for example, a lower partition wall in FIG. 7) of a pair of partition walls which partitions, respectively, the continuous discharge port and one of suction ports 21 and 31 which confronts pump chambers PR having larger internal volumes is constituted as a partition member which is a different member from the both members 11 and 12, and which is arranged to be relatively moved to the both members 11 and 12. The suction region may be enlarged by continuously varying a proportion of the suction region and the discharge region by moving this partition member in the circumferential direction.

In this way, it is possible to ensure the suction amount at the high engine speed by continuously varying the proportion of the suction region and the discharge region. Moreover, it is possible to suppress the sudden variation of the pump discharge amount by the expansion and the contraction of the suction region and the discharge region. It is possible to perform the smooth discharge control of pump 10. Furthermore, in the above-described structure, it is possible to readily expand and contract the suction region and the discharge region only by moving the partition member. With this, it is possible to improve the responsiveness of pump 10.

In the oil pumps according to the second and third embodiments of the present invention, the basic engine speed is used as the control parameters of solenoid valve SV and linear solenoid valve RSV. Moreover, in addition to the engine speed, solenoid valve SV and linear solenoid valve RSV may be controlled in accordance with, for example, the temperature of the cooling water and/or the hydraulic fluid flowing in the engine, and the pressures within switching ports 23 and 33 which are sensed by pressure sensors and pressure switches within switching ports 23 and 33.

That is, the viscosity of the hydraulic fluid affects the generation of the cavitation. Accordingly, it is possible to more appropriately suppress the generation of the cavitation by setting, as the control parameters, the temperature of the cooling water or the temperature of the hydraulic fluid, preferably the both of the temperatures of the cooling water and the hydraulic fluid. Moreover, the cavitation is generated when the internal pressures of switching ports 23 and 33 become the negative pressures. Accordingly, it is effective that the port switching control is performed based on the internal pressures of switching ports 23 and 33.

An oil pump according to the embodiment of the present invention includes: a pump constituting section arranged to continuously increase or decrease volumes of a plurality of hydraulic fluid chambers by being driven by an internal combustion engine; a port block including; a suction port opened in a region in which volumes of the hydraulic fluid chambers are increased, a switching port opened on an upstream side of a region in which the volumes of the hydraulic fluid chambers are decreased, and a discharge port opened on a downstream side of the region in which the volumes of the hydraulic fluid chambers are decreased, a valve receiving hole including; a first port formed in an inner circumference surface of the valve receiving hole, and connected with the suction port, a second port formed in the inner circumference surface of the valve receiving hole, and connected with the switching port, a third port formed in the inner circumference surface of the valve receiving hole, and connected with the discharge port, and a discharge pressure introduction port formed at a first end portion of the valve receiving hole, and arranged to receive the discharge pressure which is the hydraulic pressure within the discharge port, a valve element slidably received within the valve receiving hole, and arranged to switch a first state in which the second port and the third port are connected and a connection between the first port and the second port is restricted, and a second state in which the first port and the second port are connected and the connection between the second port and the third port is restricted; and an urging member provided on a second end side of the valve receiving hole, and arranged to urge the valve element toward the first end side of the valve receiving hole, the valve element having a pressure receiving area which is on the first end side of the valve element, and to which the discharge pressure is acted, the pressure receiving area of the valve element and the urging force of the urging member being set so that the oil pump is switched from the first state to the second state when the pressure within the hydraulic fluid chambers to which the switching port is opened is a negative pressure.

In the oil pump oil pump according to the embodiment of the present invention, the pump constituting section is arranged to vary the discharge amount in accordance with the discharge pressure; and the discharge pressure by which the valve element switches from the first state to the second state is set to a value smaller than a discharge pressure at which the discharge amount per one rotation of the oil pump is minimized.

The cavitation is generated before the discharge pressure of the pump reaches the discharge pressure at which the inherent discharge amount of the pump is minimized. Accordingly, it is possible to suppress the cavitation by setting the switching pressure to a value smaller than the discharge pressure at which the inherent discharge amount is minimized.

In the oil pump oil pump according to the embodiment of the present invention, the pump constituting section includes a rotor which is driven and rotated by the internal combustion engine, and which has a plurality of grooves that are formed on an outer circumference side, and that extend in the radial directions, a cam ring which is disposed radially outside the rotor, and which is moved so that a center of an inner circumference surface of the cam ring is eccentric from a center of the rotation of the rotor, and vanes each of which is received within one of the plurality of the grooves, each of which is moved in a radially inward direction and in a radially outward direction, and each of which has one end portion abutted and slid on the inner circumference surface of the is cam ring; and the port block is disposed on both sides of the cam ring in the axial direction so that each of the hydraulic fluid chambers is separated by the port block, the rotor, the cam ring, and the vanes.

In the oil pump oil pump according to the embodiment of the present invention, an eccentric amount between the center of the rotation of the rotor and the center of the inner circumference surface of the cam ring is decreased when the discharge pressure reaches a first pressure; the eccentric amount between the center of the rotation of the rotor and the center of the inner circumference surface of the cam ring is further decreased when the discharge pressure reaches a second pressure higher than the first pressure; and the discharge pressure at which the valve element switches from the first state to the second state is set greater than the first pressure, and lower than the second pressure.

Accordingly, it is possible to obtain the operations and the effects of the oil pump according to the embodiments of the present invention while the functions of the variable displacement pump are ensured. That is, the discharge amount is decreased by switching from the first state to the second state. The timing of the swing movement of the cam ring is delayed. With this, the pump performs unnecessary work. Consequently, when the above-described switching is performed in a state in which the cavitation cannot be generated, it is not possible to attain the effect of the decrease of the cavitation. Moreover, the functions of the variable displacement pump that decreases the driving torque of the pump is deteriorated.

In the oil pump oil pump according to the embodiment of the present invention, the oil pump further comprises two springs which have different spring loads, and which act urging forces on the cam ring; the cam ring is moved against the urging force of one of the springs having a relatively small spring load, in a direction to decrease the eccentric amount between the center of the rotation of the rotor and the center of the inner circumference surface of the cam ring, when the discharge pressure reaches the first pressure; and the cam ring is moved against the urging force of one of the springs having the relatively large spring load, in a direction to further decrease the eccentric amount between the center of the rotor and the center of the inner circumference surface of the cam ring, when the discharge pressure reaches the second pressure.

In the oil pump oil pump according to the embodiment of the present invention, a side surface of the cam ring covers a part of the suction port when the eccentric amount between the center of the rotation of the rotor and the center of the inner circumference surface of the cam ring is decreased.

Accordingly, it is possible to decrease the suction amount from the suction port, and thereby to suppress the generation of the cavitation.

In the oil pump oil pump according to the embodiment of the present invention, the pump constituting section is driven by a rotational speed larger than a rotational speed of a crank shaft of the internal combustion engine.

Accordingly, when the pumps having the same ability are compared, it is possible to decrease the size of the pump relative to the conventional pump, by increasing the rotational speed of the pump relative to the conventional pump.

In the oil pump oil pump according to the embodiment of the present invention, the pump constituting section is driven by twice the rotational speed of the crank shaft.

In this way, the rotational speed of the pump is set to twice the rotational speed of the conventional pump. With this, when the pumps having the same ability are compared, half of the volume of the conventional oil pump is sufficient for the oil pump according to the embodiments of the present invention. Accordingly, it is possible to considerably decrease the size of the pump.

In the oil pump oil pump according to the embodiment of the present invention, the pump constituting section is driven by a balancer apparatus arranged to decrease a secondary oscillation of the internal combustion engine.

In this way, the existing device provided to the internal combustion engine is used. With this, it is possible to obtain twice the rotational driving force of the conventional oil pump without using new driving force. Accordingly, it is possible to suppress the increase of the cost.

In the oil pump oil pump according to the embodiment of the present invention, the pump constituting section is assembled in a balancer apparatus.

In this way, the pump and the balancer apparatus are integrally formed. With this, it is possible to mount these integrated apparatus on the vehicle and so on, relative to a case in which the pump is mounted as the different member.

In the oil pump oil pump according to the embodiment of the present invention, the control section is electrically controlled.

Accordingly, the oil pump is not influenced by the abrasion of the oil pump, and the variation of the hydraulic pressure by the change of the kind of the oil. Consequently, it is possible to constantly perform the switching control appropriately.

In the oil pump oil pump according to the embodiment of the present invention, the control section is controlled in accordance with a sensed engine speed of the internal combustion engine.

The generation of the cavitation depends on the rotational speed of the pump. Accordingly, it is possible to appropriately suppress the generation of the cavitation by the control according to the engine speed relating to the rotational speed of the pump.

In the oil pump oil pump according to the embodiment of the present invention, the control section is controlled in accordance with a sensed temperature of the internal combustion engine.

The generation of the cavitation depends on the viscosity of the hydraulic fluid. By the above described structure, it is possible to more appropriately suppress the generation of the cavitation.

In the oil pump oil pump according to the embodiment of the present invention, the control section is controlled in accordance with a sensed pressure of the switching port.

When the pressure within the switching port opened on the upstream side of the region in which the volumes is decreased becomes the negative pressure, the cavitation are generated. Accordingly, it is possible to appropriately suppress the generation of the cavitation by the switching control based on the pressure of the switching port.

In the oil pump oil pump according to the embodiment of the present invention, a region in which the hydraulic fluid is sucked is enlarged and a region in which the hydraulic fluid is discharged is decreased at a rotational speed at which the hydraulic chambers on the upstream side of the region in which the volumes are decreased becomes the negative pressure.

Accordingly, it is possible to ensure the sufficient suction amount in the suction process, and thereby to suppress the cavitation.

In the oil pump oil pump according to the embodiment of the present invention, the region in which the hydraulic fluid is sucked is continuously enlarged in accordance with the rotational speed.

Accordingly, it is possible to suppress the sudden variation of the discharge amount based on the expansion and the contraction of the suction/discharge regions.

In the oil pump oil pump according to the embodiment of the present invention, the port block is provided with a partition member partitioning the suction port and the discharge port; and a suction region of the hydraulic fluid is enlarged by moving the partition member in accordance with the rotational speed.

Accordingly, it is possible to easily expand and contract the suction/discharge regions, relative to the conventional pump in which the suction/discharge regions are partitioned in the fixed manner. Consequently, it is possible to improve the responsiveness of the pump.

The entire contents of Japanese Patent Application No. 2011-31691 filed Feb. 17, 2011 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. An oil pump comprising:

a pump constituting section arranged to continuously increase or decrease volumes of a plurality of hydraulic fluid chambers by being driven by an internal combustion engine;

a port block including; a suction port opened in a region in which volumes of the hydraulic fluid chambers are increased, a switching port opened on an upstream side of a region in which the volumes of the hydraulic fluid chambers are decreased, and a discharge port opened on a downstream side of the region in which the volumes of the hydraulic fluid chambers are decreased,

a valve receiving hole including; a first port formed in an inner circumference surface of the valve receiving hole, and connected with the suction port, a second port formed in the inner circumference surface of the valve receiving hole, and connected with the switching port, a third port formed in the inner circumference surface of the valve receiving hole, and connected with the discharge port, and a discharge pressure introduction port formed at a first end portion of the valve receiving hole, and arranged to receive the discharge pressure which is the hydraulic pressure within the discharge port,

a valve element slidably received within the valve receiving hole, and arranged to switch a first state in which the second port and the third port are connected and a connection between the first port and the second port is restricted, and a second state in which the first port and the second port are connected and the connection between the second port and the third port is restricted; and

an urging member provided on a second end side of the valve receiving hole, and arranged to urge the valve element toward the first end side of the valve receiving hole,

the valve element having a pressure receiving area which is on the first end side of the valve element, and to which the discharge pressure is acted, the pressure receiving area of the valve element and the urging force of the urging member being set so that the oil pump is switched from the first state to the second state when the pressure within the hydraulic fluid chambers to which the switching port is opened is a negative pressure.

2. The oil pump as claimed in claim 1, wherein the pump constituting section is arranged to vary the discharge amount in accordance with the discharge pressure; and the discharge pressure by which the valve element switches from the first state to the second state is set to a value smaller than a discharge pressure at which the discharge amount per one rotation of the oil pump is minimized.

3. The oil pump as claimed in claim 2, wherein the pump constituting section includes a rotor which is driven and rotated by the internal combustion engine, and which has a plurality of grooves that are formed on an outer circumference side, and that extend in the radial directions, a cam ring which is disposed radially outside the rotor, and which is moved so that a center of an inner circumference surface of the cam ring is eccentric from a center of the rotation of the rotor, and vanes each of which is received within one of the plurality of the grooves, each of which is moved in a radially inward direction and in a radially outward direction, and each of which has one end portion abutted and slid on the inner circumference surface of the cam ring; and the port block is disposed on both sides of the cam ring in the axial direction so that each of the hydraulic fluid chambers is separated by the port block, the rotor, the cam ring, and the vanes.

4. The oil pump as claimed in claim 3, wherein an eccentric amount between the center of the rotation of the rotor and the center of the inner circumference surface of the cam ring is decreased when the discharge pressure reaches a first pressure; the eccentric amount between the center of the rotation of the rotor and the center of the inner circumference surface of the cam ring is further decreased when the discharge pressure reaches a second pressure higher than the first pressure; and the discharge pressure at which the valve element switches from the first state to the second state is set greater than the first pressure, and lower than the second pressure.

5. The oil pump as claimed in claim 4, wherein the oil pump further comprises two springs which have different spring loads, and which act urging forces on the cam ring; the cam ring is moved against the urging force of one of the springs having a relatively small spring load, in a direction to decrease the eccentric amount between the center of the rotation of the rotor and the center of the inner circumference surface of the cam ring, when the discharge pressure reaches the first pressure; and the cam ring is moved against the urging force of one of the springs having the relatively large spring load, in a direction to further decrease the eccentric amount between the center of the rotor and the center of the inner circumference surface of the cam ring, when the discharge pressure reaches the second pressure.

6. The oil pump as claimed in claim 4, wherein a side surface of the cam ring covers a part of the suction port when the eccentric amount between the center of the rotation of the rotor and the center of the inner circumference surface of the cam ring is decreased.

7. The oil pump as claimed in claim 4, wherein the pump constituting section is driven by a rotational speed larger than a rotational speed of a crank shaft of the internal combustion engine.

8. The oil pump as claimed in claim 7, wherein the pump constituting section is driven by twice the rotational speed of the crank shaft.

9. The oil pump as claimed in claim 8, wherein the pump constituting section is driven by a balancer apparatus arranged to decrease a secondary oscillation of the internal combustion engine.

10. The oil pump as claimed in claim 1, wherein the pump constituting section is assembled in a balancer apparatus.

11. An oil pump comprising:

a pump constituting section arranged to continuously increase or decrease volumes of a plurality of hydraulic fluid chambers by being driven by an internal combustion engine;

a port block including; a suction port opened in a region in which volumes of the hydraulic fluid chambers are increased, a switching port opened on an upstream side of a region in which the volumes of the hydraulic fluid chambers are decreased, and a discharge port opened on a downstream side of the region in which the volumes of the hydraulic fluid chambers are decreased, and

a control section configured to control so that the switching port discharges the hydraulic fluid to an outside with the discharge port at a low rotational speed at which the pressure within the hydraulic fluid chamber to which the switching port is opened is a positive pressure, and to control so that the hydraulic fluid is supplied from a low pressure portion connected with the suction port, to the switching port at a high rotational speed at which the pressure within the hydraulic fluid chambers to which the switching port is opened is a negative pressure.

12. The oil pump as claimed in claim 11, wherein the control section is electrically controlled.

13. The oil pump as claimed in claim 12, wherein the control section is controlled in accordance with a sensed engine speed of the internal combustion engine.

14. The oil pump as claimed in claim 13, wherein the control section is controlled in accordance with a sensed temperature of the internal combustion engine.

15. The oil pump as claimed in claim 12, wherein the control section is controlled in accordance with a sensed pressure of the switching port.

16. An oil pump comprising:

a pump constituting section arranged to continuously increase or decrease volumes of a plurality of hydraulic fluid chambers by being driven by an internal combustion engine;

a port block including; a suction port opened to at least a region in which volumes of the hydraulic fluid chambers are increased, and a discharge port opened to a region in which volumes of the hydraulic fluid chambers are decreased; and

a control section configured to control so that, at a rotational speed at which the pressure within the hydraulic fluid chambers on an upstream side of the region in which the volumes of the hydraulic fluid chambers are decreased becomes a negative pressure, the hydraulic fluid is supplied from a low pressure portion to the hydraulic fluid chambers which are the negative pressure.

17. The oil pump as claimed in claim 16, wherein a region in which the hydraulic fluid is sucked is enlarged and a region in which the hydraulic fluid is discharged is decreased at a rotational speed at which the hydraulic chambers on the upstream side of the region in which the volumes are decreased becomes the negative pressure.

18. The oil pump as claimed in claim 17, wherein the region in which the hydraulic fluid is sucked is continuously enlarged in accordance with the rotational speed.

19. The oil pump as claimed in claim 18, wherein the port block is provided with a partition member partitioning the suction port and the discharge port; and a suction region of the hydraulic fluid is enlarged by moving the partition member in accordance with the rotational speed.