JP2013057326A - Variable displacement pump - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a variable displacement oil pump useful for miniaturization of a pump.SOLUTION: This variable displacement oil pump 10 is a vane type variable displacement oil pump varying the discharge amount by operation of a cam ring 17. A first pressure chamber 31 in which a discharge pressure is introduced and which includes a first pressure receiving surface 33 making the discharge pressure act on a cam ring 17 in a direction to reduce an eccentricity amount, and a second pressure chamber 32 which includes a second pressure receiving surface 34 making the discharge pressure act on the cam ring 17 in a direction to increase the eccentricity amount are oppositely arranged at an outer peripheral area of the cam ring 17 at a discharge area side with a pivot section 17a put therebetween. The eccentricity quantity of the cam ring 17 is controlled by a balance of inner pressures of both pressure chambers 31, 32.

Description

  The present invention is applied to, for example, a hydraulic source that supplies hydraulic oil to each sliding portion of an internal combustion engine for automobiles, and in particular, a variable displacement pump that can vary a discharge amount (discharge pressure) according to an operating state. About.

  As a conventional variable displacement oil pump applied to an oil pump for automobiles, for example, the one described in Patent Document 1 below is known.

  Briefly, this variable displacement pump is a so-called vane pump that selectively supplies discharge pressure to two pressure chambers defined between a housing and a cam ring, and rotates the rotor by a spring. The discharge amount (discharge pressure) is made variable by controlling the amount of eccentricity of the cam ring urged in a direction that is always eccentric with respect to the center.

Special table 2008-524500 gazette

  However, the conventional variable displacement oil pump has a structure in which the biasing force of the spring is balanced against the oil pressure caused by the internal pressure (discharge pressure) of the two pressure chambers. As a result, there is a problem that the pump must be enlarged.

  The present invention has been devised in view of the technical problem of the conventional variable displacement oil pump, and an object thereof is to provide a variable displacement oil pump for use in reducing the size of the pump.

  The present invention is a so-called vane type variable displacement oil pump, which is defined in particular on the outer peripheral side of the cam ring, and in which discharge pressure is introduced into the cam ring to reduce the amount of eccentricity with respect to the cam ring. A first pressure chamber having a first pressure receiving surface on which the discharge pressure is applied and a second pressure chamber having a second pressure receiving surface on which the discharge pressure is applied in a direction to increase the eccentric amount with respect to the cam ring are provided. .

  According to the present invention, since the amount of eccentricity of the cam ring is controlled by balancing the internal pressures of the two pressure chambers, it is not always necessary to use a biasing member such as a spring. Even if it is used, it is not necessary to increase the urging force of the urging member, so that the pump can be reduced in size.

1 is an exploded perspective view showing a configuration of a variable displacement oil pump according to a first embodiment of the present invention. It is a front view which shows the state which removed the cover member about the variable displacement oil pump which concerns on the same embodiment, Comprising: The amount of eccentricity of a cam ring represents the maximum state. It is a front view which shows the state which removed the cover member about the variable displacement oil pump which concerns on the same embodiment, Comprising: The eccentric amount of a cam ring represents the state with the minimum. It is the sectional view on the AA line of FIG. It is a front view which shows the inside of the housing which concerns on the same embodiment. It is a longitudinal cross-sectional view which shows the state at the time of the non-energization of the solenoid valve which concerns on the same embodiment. It is a longitudinal cross-sectional view which shows the state at the time of energization of the solenoid valve which concerns on the same embodiment. FIG. 3 is a hydraulic circuit diagram of the variable displacement oil pump according to the same embodiment. It is a graph which shows the relationship between the rotation speed of an engine, and oil_pressure | hydraulic about the variable displacement oil pump which concerns on the same embodiment. It is a longitudinal cross-sectional view which shows the state at the time of the non-energization of the solenoid valve which concerns on the modification of 1st Embodiment of this invention. It is a longitudinal section showing the state at the time of energization of the solenoid valve concerning the modification. It is a disassembled perspective view which shows the structure of the variable displacement oil pump which concerns on 2nd Embodiment of this invention. It is a front view which shows the state which removed the cover member about the variable displacement oil pump which concerns on the same embodiment, Comprising: The amount of eccentricity of a cam ring represents the maximum state. It is a front view which shows the state which removed the cover member about the variable displacement oil pump which concerns on the same embodiment, Comprising: The eccentric amount of a cam ring represents the state with the minimum. It is a front view of the cover member concerning a 3rd embodiment of the present invention. It is a rear view of the cover member concerning the embodiment. It is a front view which shows the state which removed the cover member about the variable displacement oil pump which concerns on 4th Embodiment of this invention, Comprising: The amount of eccentricity of a cam ring represents the maximum state. It is a front view which shows the state which removed the cover member about the variable displacement oil pump which concerns on 4th Embodiment of this invention, Comprising: The eccentric amount of a cam ring represents the state with the minimum. It is a longitudinal cross-sectional view which shows the non-operation state of the hydraulic direction switching valve concerning 5th Embodiment of this invention. It is a longitudinal cross-sectional view which shows the operating state of the hydraulic direction switching valve concerning the embodiment. FIG. 3 is a hydraulic circuit diagram of the variable displacement oil pump according to the same embodiment. It is a graph which shows the relationship between the rotation speed of an engine and oil_pressure | hydraulic about the variable displacement pump which concerns on the same embodiment.

  Embodiments of a variable displacement oil pump according to the present invention will be described below in detail with reference to the drawings. In each of the embodiments, the variable displacement oil pump is applied as an oil pump that supplies engine lubricating oil to a valve timing control device that controls the opening / closing timing of the sliding portion of the automobile internal combustion engine and the engine valve. An example is shown.

  1 to 9 show a first embodiment of a variable displacement oil pump according to the present invention. This oil pump 10 is provided at the front end of a cylinder block of an internal combustion engine, etc. As shown in FIG. 1, a pump body 11 having a U-shaped cross section having a pump accommodating chamber 13 formed in a cylindrical space inside and having a one end side opened, and a cover member 12 that closes one end opening of the pump body 11. A drive shaft 14 that is rotatably supported by the housing and is driven to rotate by the crankshaft of the engine through the substantially central portion of the pump housing chamber 13, and is housed rotatably in the pump housing chamber 13. Thus, the central portion of the rotor 15 is coupled to the drive shaft 14 and the plurality of slits 15a radially formed in the outer peripheral portion of the rotor 15 are housed in a freely retractable manner. And a pump chamber 20 which is a plurality of hydraulic fluid chambers together with the rotor 15 and the adjacent vanes 16 and 16. , A spring 18 that is housed in the pump body 11 and is a biasing member that constantly biases the cam ring 17 in a direction in which the eccentric amount of the cam ring 17 with respect to the rotation center of the rotor 15 increases. And a pair of ring members 19, 19 having a smaller diameter than the rotor 15 slidably disposed on both side portions on the inner peripheral side.

  The pump body 11 is integrally formed of an aluminum alloy material. As shown in FIGS. 4 and 5, one end portion of the drive shaft 14 is rotated at a substantially central position of the bottom surface 13 a of the pump housing chamber 13. A bearing hole 11a that is freely supported is formed through. Further, as shown in FIG. 5, a support groove 11b having a substantially semicircular cross section for supporting the cam ring 17 in a swingable manner is cut out at a predetermined position on the inner peripheral wall of the pump housing chamber 13 which is the inner surface of the pump body 11. Is formed. Further, on the inner peripheral wall of the pump housing chamber 13, on both sides of a straight line connecting the center of the bearing hole 11 a and the center of the support groove 11 b (hereinafter referred to as “cam ring reference line”) M, on the outer peripheral portion of the cam ring 17. First and second seal slidable contact surfaces 11c and 11d are formed so that seal members 30 and 30 described later are in slidable contact with each other. Each of the seal sliding contact surfaces 11c and 11d has an arcuate surface shape constituted by predetermined radii R1 and R2 from the center of the support groove 11b, and the seal member 30 in the eccentric swing range of the cam ring 17. , 30 are set to circumferential lengths that can always slide. As a result, when the cam ring 17 is eccentrically oscillated, the cam ring 17 is slidably guided along the seal sliding contact surfaces 11c and 11d, so that the cam ring 17 can be smoothly operated (eccentric oscillation). It is like that.

  Further, as shown in FIGS. 2 and 5, the bottom surface 13 a of the pump storage chamber 13 is a region where the internal volume of the pump chamber 20 increases in the outer peripheral region of the bearing hole 11 a due to the pump action of the pump element. The suction port 21, which is a substantially arc-shaped suction portion that opens to the (suction region), opens to a region (discharge region) in which the internal volume of the pump chamber 20 decreases with the pump action of the pump element. The discharge port 22 which is a substantially arc-shaped discharge part is cut out so as to be substantially opposed to each other across the bearing hole 11a.

  The suction port 21 is connected to an introduction passage 24 extending from the substantially central position of the suction port 21 to the later-described spring accommodating chamber 28, and the bottom wall of the pump body 11 is placed in the middle of the introduction passage 24. A suction hole 21a that penetrates and opens to the outside is formed. As a result, as shown in FIG. 8, the lubricating oil stored in the oil pan 52 of the engine is sucked through the suction hole 21a and the suction port 21 based on the negative pressure generated by the pump action of the pump element. Each pump chamber 20 in the region is sucked. The suction hole 21a is configured so as to face the outer peripheral region of the cam ring 17 on the pump suction side together with the introduction passage 24, and guides the suction pressure to the outer peripheral region of the cam ring 17 on the pump suction side. Yes. As a result, the outer peripheral area of the cam ring 17 on the pump suction side adjacent to each pump chamber 20 in the suction area becomes suction pressure or atmospheric pressure, and therefore, the cam ring 17 on the pump suction side from each pump chamber 20 in the suction area. This is used for suppressing leakage of lubricating oil to the outer peripheral region. Here, the pump suction side means a region on the left side of a cam ring eccentric direction line N described later in FIG.

  The discharge port 22 is connected to an introduction passage 25 extending from the starting end of the discharge port 22 to the first pressure chamber 31 described later, which is defined on the outer peripheral side of the cam ring 17. A discharge hole 22a that penetrates the bottom wall of the pump body 11 and opens to the outside is formed through the end portion. The discharge hole 22a is connected to each sliding portion in the engine and a valve timing control device via an oil main gallery not shown. With this configuration, the lubricating oil pressurized by the pumping action of the pump element and discharged from the pump chambers 20 in the discharge region passes through the discharge ports 22 and the discharge holes 22a. It is supplied to the timing control device. The discharge hole 22a is configured so as to face the outer peripheral region of the cam ring 17 on the pump discharge side together with the introduction passage 25, and guides the discharge pressure to the outer peripheral region of the cam ring 17 on the pump discharge side. Yes. Here, the pump discharge side means a region on the right side of a cam ring eccentric direction line N described later in FIG.

  Further, a communication groove 23 that communicates the discharge port 22 and the bearing hole 11a is formed in the vicinity of the starting end portion of the discharge port 22, and lubricating oil is supplied to the bearing hole 11a through the communication groove 23. In addition to the supply, lubricating oil is also supplied to the sides of the rotor 15 and the vane 16 to ensure the lubricity of each sliding portion. The communication groove 23 is formed so as not to coincide with the direction in which the vanes 16 are projected and retracted, and the dropout to the communication groove 23 when the vanes 16 appear and disappear is suppressed.

  The cover member 12 has a substantially plate shape, and a position corresponding to the bearing hole 11a of the pump body 11 on the outer side is formed to be slightly thicker. A bearing hole 12a for rotatably supporting the shaft is formed. In the present embodiment, the inner surface of the cover member 12 is substantially flat. However, the suction and discharge ports 21 and 22 can be formed in the same manner as the bottom surface of the pump body 11. It is also possible to provide a groove for guiding the lubricating oil to the bearing hole 12 a such as the communication groove 23. The cover member 12 is attached to the opening end surface of the pump body 11 with a plurality of bolts 26.

  The drive shaft 14 is configured to rotate the rotor 15 in the clockwise direction in FIG. 2 by the rotational force transmitted from the crankshaft, and a straight line (hereinafter referred to as the straight line M) at the center of the drive shaft 14. 2) With N as a boundary, the left half in FIG. 2 is the pump suction side and the right half is the pump discharge side.

  As shown in FIGS. 1 and 2, the rotor 15 has a plurality of slits 15a radially formed radially outward from the inner center side, and an inner base end portion of each slit 15a. Are formed with a back pressure chamber 15b having a substantially circular cross section for introducing the discharge oil discharged to the discharge port 22, respectively. Thereby, each vane 16 is pushed outward by the centrifugal force accompanying the rotation of the rotor 15 and the hydraulic pressure of the back pressure chamber 15b.

  Each vane 16 has a distal end surface in sliding contact with the inner peripheral surface of the cam ring 17 and a side surface of each base end portion in sliding contact with the outer sliding surfaces of the ring members 19, 19. As a result, even when the engine speed is low and the centrifugal force or the hydraulic pressure of the back pressure chamber 15b is small, the outer peripheral surface of the rotor 15, the inner surfaces of the adjacent vanes 16, 16 and the inner peripheral surface of the cam ring 17, and the side wall The bottom surface 13a of the pump housing chamber 13 of the pump body 11 and the inner side surface of the cover member 12 define each pump chamber 20 in a liquid-tight manner.

  The cam ring 17 is integrally formed of a so-called sintered metal in a substantially cylindrical shape, and is fitted in the support groove 11b of the pump body 11 at a predetermined position on the outer peripheral portion so as to constitute an eccentric rocking fulcrum. 17a protrudes along the axial direction, and an arm portion 17b linked to the spring 18 protrudes along the axial direction at a position opposite to the pivot portion 17a across the center of the cam ring 17. ing.

  Here, in the pump body 11, a spring accommodating chamber 28 is disposed at a position opposite to the support groove 11 b so as to communicate with the pump accommodating chamber 13 via a communicating portion 27 set to a predetermined width L. The spring 18 is accommodated in the spring accommodating chamber 28. The spring 18 is elastically held with a predetermined set load W between the lower surface of the distal end portion of the arm portion 17 b that extends into the spring accommodating chamber 28 through the communication portion 27 and the bottom surface of the spring accommodating chamber 28. Yes. A support protrusion 17i formed in a substantially arc shape that engages with the inner peripheral side of the spring 18 protrudes from the lower surface of the distal end of the arm portion 17b, and one end of the spring 18 is projected by the support protrusion 17i. Is supported.

  With this configuration, the spring 18 constantly biases the cam ring 17 in the direction in which the eccentric amount increases (clockwise in FIG. 2) via the arm portion 17b with an elastic force based on the set load W. It is like that. Thus, in the non-operating state of the cam ring 17 shown in FIG. 2, the cam ring 17 is pressed against the stopper portion 28 a, the upper surface of the arm portion 17 b projecting from the lid portion of the spring accommodating chamber 28 by the urging force of the spring 18. And is regulated to a position where the amount of eccentricity is maximized. In this way, the arm portion 17b is extended to the opposite side of the pivot portion 17a, and the tip end portion of the arm portion 17b is urged by the spring 18, so that the cam ring 17 is maximized. Since torque can be generated, the spring 18 can be reduced in size, and as a result, the pump itself can be reduced in size.

  Further, on the outer peripheral portion of the cam ring 17, the first and second seal sliding contact surfaces 11c, 11d formed so as to face the first and second sealing sliding contact surfaces 11c, 11d are concentric arc-shaped first, A pair of first and second seal constituent parts 17c, 17d having a substantially triangular cross section having the second seal surfaces 17g, 17h are projected along the axial direction, and the respective seal constituent parts 17c, 17d. First and second seal holding grooves 17e and 17f having a substantially rectangular cross section are formed in the seal surfaces 17g and 17h along the axial direction, and cam ring 17 is formed in each of the seal holding grooves 17e and 17f. A pair of seal members 30, 30 that are in sliding contact with the seal sliding contact surfaces 11c, 11d during eccentric swinging are respectively housed and held.

  Here, the seal surfaces 17g and 17h are formed at a predetermined radius R3 slightly smaller than the radii R1 and R2 constituting the seal sliding contact surfaces 11c and 11d corresponding to the center of the pivot portion 17a. A small clearance C is formed between the seal surfaces 17g and 17h and the seal sliding contact surfaces 11c and 11d.

  Each of the seal members 30 and 30 is, for example, a long and narrow line formed along the axial direction of the cam ring 17 with a fluorine-based resin material having low friction characteristics, and is disposed at the bottom of each of the seal holding grooves 17e and 17f. The elastic members 29 and 29 are made to be pressed against the seal sliding contact surfaces 11c and 11d by the elastic force. Thereby, the good liquid tightness of each pressure chamber 31 and 32 mentioned later is always ensured.

  When the cam ring 17 is in an inoperative state, an outer peripheral surface of the cam ring 17 and an inner surface of the pump body 11 are located on the outer peripheral area of the cam ring 17 on the pivot portion 17a side of the cam ring eccentric direction line N on the pump discharge side. Between the first pressure chamber 31 and the second pressure chamber 32 on both sides of the pivot portion 17a with the outer peripheral surface of the cam ring 17, the pivot portion 17a, the inner surfaces of the seal members 30 and 30 and the pump body 11. Are each defined. In the present embodiment, the first and second pressure chambers 31 and 32 are shown as examples in the outer peripheral area of the cam ring 17 within the range on the pump discharge side. It is better to store in a region that overlaps with the discharge region that becomes the pressurizing region in the radial direction, that is, in a region facing the pump chamber 20 that is always positive pressure across the peripheral wall of the cam ring 17.

  The discharge pressure discharged to the discharge port 22 is always introduced into the first pressure chamber 31 through the introduction passage 25, and the outer peripheral surface of the cam ring 17 facing the first pressure chamber 31 is used. A direction in which the eccentric amount is reduced with respect to the cam ring 17 (counterclockwise direction in FIG. 2) by applying a discharge pressure to the first pressure receiving surface 33 that is configured and receives a force that acts to prevent the biasing force of the spring 18. ) Is given a swinging force (moving force). That is, the first pressure chamber 31 constantly biases the cam ring 17 in the direction in which the center of the cam ring 17 approaches the concentricity with the rotation center of the rotor 15 via the first pressure receiving surface 33. This is used to control the amount of movement in the concentric direction.

  On the other hand, the second pressure chamber 32 is formed through the bottom wall of the pump body 11 and is connected to the discharge hole 22a via a solenoid valve 40 described later that is controlled according to the engine operating state. The discharge pressure is appropriately introduced through the second pressure receiving member, and is constituted by the outer peripheral surface of the cam ring 17 facing the second pressure chamber 32 and receives a force acting in the direction of assisting the biasing force of the spring 18. By applying a discharge pressure to the surface 34, a swinging force is applied to the cam ring 17 in a direction that increases the amount of eccentricity (clockwise in FIG. 2).

  Here, as shown in FIG. 2, the pressure receiving area S <b> 2 of the second pressure receiving surface 34 is set smaller than the pressure receiving area S <b> 1 of the first pressure receiving surface 33, and is applied based on the internal pressure of the second pressure chamber 32. The urging force in the eccentric direction of the cam ring 17 by the urging force and the urging force of the spring 18 and the urging force by the first pressure chamber 31 are configured to be balanced with a predetermined force relationship. The urging force due to the pressure assists the urging force of the spring 18. That is, the second pressure chamber 32 causes the cam ring 17 to appropriately assist the urging force of the spring 18 by causing the discharge pressure supplied as needed via the solenoid valve 40 to act on the second pressure receiving surface 34. This is used to control the amount of movement in the eccentric direction.

  Further, as shown in FIG. 8, the oil pump 10 is provided with a solenoid valve 40 that operates according to the operating state of the engine based on the excitation current from the vehicle-mounted ECU 51, separately from the oil pump 10. The discharge hole 22a and the introduction hole 35 are connected via the solenoid valve 40, whereby the first pressure chamber 31 and the second pressure chamber 32 communicate with each other when the solenoid valve 40 is opened.

  As shown in FIGS. 6 and 7, the solenoid valve 40 includes a cylindrical valve body 41 that is open at one end and closed at the other end, and is slidably accommodated in the valve body 41 in the axial direction. The valve body 42 is formed with first and second land portions 42a and 42b that are in sliding contact with the inner peripheral surface of the valve body 41 at both ends, and the second land portion 42b of the valve body 42 A spring 43 that is housed in a back pressure chamber 45 defined on the end side and urges the valve body 42 toward one end side of the valve body 41, and is attached to the opening end of the valve body 41. And an electromagnetic unit 44 that moves the valve body 42 in the axial direction to the other end side of the valve body 41 against the biasing force of the spring 43.

  The valve body 41 has, on its peripheral wall, an IN port 41a connected to the discharge hole 22a, an OUT port 41b connected to the introduction hole 35, and a drain port 41c connected to the suction port 21 or the outside. A back pressure port 41d that is connected to the suction port 21 or the outside and always opens to the back pressure chamber 45 is formed through the side wall at the other end.

  The valve body 42 is formed with a reduced diameter at an intermediate portion in the axial direction, and an annular space 46 is defined between the valve body 41 and the land portions 42a and 42b. The port 41b communicates with the IN port 41a or the drain port 41c.

  The electromagnetic unit 44 is configured in a well-known manner, a coil is wound around a bobbin, and a yoke is fitted around the coil unit 44a. The electromagnetic unit 44 can be moved back and forth in the axial direction on the inner peripheral side of the coil unit 44a. It is mainly composed of an armature (not shown) made of a magnetic material provided, and a rod 44b coupled to the armature and moving forward and backward with the armature according to the energized state.

  Here, as shown in FIG. 6, the solenoid valve 40 is configured as a so-called normally open type, and therefore, in a non-energized state, the IN port 41a and the OUT port 41b communicate with each other through the annular space 56. Thus, the discharge pressure is introduced into the second pressure chamber 32 (first state according to the present invention). At this time, the drain port 41 c is open to the back pressure chamber 45.

  On the other hand, as shown in FIG. 7, when the exciting current is applied to the coil unit 44 a, the valve body 42 moves to the other end side of the valve body 41 against the urging force of the spring 43 by the pressing force of the rod 44 b. It will be pushed back. As a result, the IN port 41a is blocked by the first land portion 42a of the valve body 42, the OUT port 41b communicates with the drain port 41c through the annular space 46, and the second pressure chamber 32 has the suction pressure or It will be open to atmospheric pressure (second state according to the present invention).

  With the above configuration, the oil pump 10 acts on the cam ring 17 of the internal pressure of the first pressure chamber 31 and the urging force of the spring 18 and the internal pressure of the second pressure chamber 32 controlled by the solenoid valve 40. The amount of eccentricity of the cam ring 17 is controlled by controlling the relative force relationship. Then, by controlling the amount of eccentricity and controlling the amount of change in the internal volume of each pump chamber 20 during the pump action, the discharge pressure characteristic of the oil pump 10 is controlled.

  Hereinafter, the characteristic action of the oil pump 10 according to the present embodiment, that is, the discharge pressure control of the pump based on the eccentric amount control of the cam ring 17 will be described based on FIG. 2, FIG. 3, and FIG.

  First, the discharge pressure of the oil pump 10 is determined by the required oil pressure in each sliding part of the engine and the valve timing control device. Therefore, since the required oil pressure of the engine changes depending on the operating state of the engine, the required oil pressure varies, and a representative one is shown in the map of FIG. That is, for example, when a valve timing control device is used for the purpose of improving fuel efficiency, the required oil pressure is P1 in the figure. The required oil pressure of the internal combustion engine is determined mainly by the oil pressure required for lubricating the crankshaft bearing. This required oil pressure varies depending on the engine speed, load (throttle opening), oil temperature, etc. For example, when the load is low or when the oil temperature is low, P2 in the figure is the required oil pressure, and when the load is high or the oil temperature is high, P4 is the required oil pressure. Further, when the engine is under a high load, it is necessary to use an oil jet for cooling the piston, and the hydraulic pressure P3 in the drawing is required at a predetermined rotational speed n which is a middle rotation.

  Therefore, the oil pump 10 has a low pressure characteristic X, which is a first discharge pressure characteristic that satisfies the required oil pressure of either P1 or P2 in FIG. 9 or both P1 and P2 when the load is low or the oil temperature is low. High pressure characteristic Y which is a second discharge pressure characteristic that satisfies the required oil pressure of either P3 or P4 in FIG. 9 or both P3 and P4 at high load or high oil temperature. By switching ON / OFF of the solenoid valve 40, the operating characteristics of the cam ring 17, that is, the first and second operating pressures Px that are discharge pressures necessary for the operation of the cam ring 17 are set. , Py are changed, and an optimum hydraulic characteristic is selected from the two hydraulic characteristics X and Y according to the operating state of the engine, so that each required hydraulic pressure of the engine is satisfied.

  In the present embodiment, as shown in FIG. 9, for the low pressure characteristic X, the required hydraulic pressure P1 of the variable valve timing control device and the required hydraulic pressure P2 at the time of high engine rotation in a low load or low oil temperature state. On the other hand, the high pressure characteristic Y is set to the hydraulic characteristic indicated by the connected broken line, and the required hydraulic pressure P3 at the middle rotation of the engine in a high load or high oil temperature state and the required hydraulic pressure at the high rotation of the engine in the state. It is set to the hydraulic characteristic indicated by the solid line connecting P4.

  That is, in the oil pump 10, the set load W of the spring 18 is set to the first operating oil pressure Px, and when the load is low or the oil temperature is low, an excitation current is passed from the ECU 51 to the solenoid valve 40. The port 41 a is shut off, and the discharge pressure is introduced only into the first pressure chamber 31. As a result, the eccentric amount of the cam ring 17 is maintained in a maximum state until the internal pressure of the first pressure chamber 31 reaches the first operating oil pressure Px (see FIG. 2), and the discharge pressure increases as the engine speed increases. It will stand up suddenly. When the internal pressure of the first pressure chamber 31 reaches the first operating oil pressure Px due to the increase in the discharge pressure, the cam ring 17 is located below the cam ring eccentric direction line N with the pivot portion 17a serving as a fulcrum. It will swing in the decreasing direction (see FIG. 3). As a result, the volume change amount of each pump chamber 20 at the time of the pump action is reduced, and as a result, the increase in the discharge pressure accompanying the increase in the engine speed is moderated. Therefore, the low pressure characteristic X shown in FIG. Will be obtained.

  Subsequently, when the low load or low oil temperature state is shifted to the high load or high oil temperature state, the excitation current to the solenoid valve 40 from the ECU 51 is cut off, and the IN port 41a and the OUT port 41b are switched. As a result, the discharge pressure is introduced not only into the first pressure chamber 31 but also into the second pressure chamber 32. Then, since the pressure acting on the second pressure receiving surface 34 of the second pressure chamber 32 works to assist the urging force of the spring 18, the internal pressure of the first pressure chamber 31 becomes the first operating oil pressure Px in FIG. Even if the pressure reaches the cam ring 17, the oil pressure difference acting on the first pressure receiving surface 33 and the second pressure receiving surface 34 due to the internal pressure of the first pressure chamber 31 and the internal pressure of the second pressure chamber 32 becomes the urging force of the spring 18. Until it reaches, the cam ring 17 is held in a state where the eccentric amount of the cam ring 17 is maximized (see FIG. 2). That is, at the time of the high load or high oil temperature, the discharge pressure is applied to the first pressure receiving surface 33 and the second pressure receiving surface 34 by the internal pressure of the first pressure chamber 31 and the internal pressure of the second pressure chamber 32 as shown in FIG. The eccentric amount of the cam ring 17 is maintained in a maximum state until the second operating oil pressure Py at which the acting oil pressure difference becomes equal to the urging force of the spring 18, and the discharge pressure rises greatly as the engine speed increases. It will be. When the internal pressure of the first pressure chamber 31 reaches the second operating oil pressure Py, the cam ring 17 swings in the direction in which the eccentric amount decreases (see FIG. 3). As a result, the volume change amount of each pump chamber 20 during the operation of the pump is reduced, and the increase in the discharge pressure accompanying the increase in the engine speed is moderated. Therefore, a high pressure characteristic Y as shown in FIG. 9 is obtained. Will be.

  Thus, in principle, in the oil pump 10, the pump discharge pressure characteristic shifts to the high pressure characteristic Y when the ECU 51 determines that a high pressure is necessary based on the engine speed, load, oil temperature, and the like. It becomes. Therefore, normally, when the engine load or oil temperature is high, the high-pressure characteristic Y is shifted. Therefore, in the above description, the engine load or oil temperature is high when the high-pressure characteristic Y is exhibited. Although the state has been described as an example, for example, the valve timing control device may require a hydraulic pressure higher than the required hydraulic pressure P1, and in such a case, the ECU 51 adjusts the solenoid in accordance with the operation signal of the valve timing control device. The valve 40 is switched, and the pump discharge pressure characteristic shifts to the high pressure characteristic Y even when the engine load, oil temperature, and the like are low. In other words, in this embodiment, the required hydraulic pressure P1 is set to the normal required hydraulic pressure of the valve timing control device, but the required hydraulic pressure P1 is controlled by valve timing according to the specifications of the vehicle to be mounted. It is also possible to set the minimum required oil pressure in the apparatus.

  Further, when the high load or high oil temperature state shifts again to the low load or low oil temperature state, the excitation current is supplied again from the ECU 51, and the solenoid valve 40 becomes in the energized state as shown in FIG. The second pressure chamber 32 is opened to atmospheric pressure or suction pressure. As a result, the operation of the cam ring 17 depends on the force relationship between the internal pressure of the first pressure chamber 31 and the urging force of the spring 18, and the discharge pressure characteristic of the pump is changed to the low pressure characteristic X. As a result, it is possible to reduce the discharge pressure that becomes unnecessary due to the shift to the low load or low oil temperature state, and to suppress the power loss of the engine.

  In this way, the oil pump 10 changes the operating characteristics of the cam ring 17 by the ECU 51 switching the solenoid valve 40 based on various operating information such as the engine speed, load, oil temperature, etc. The discharge pressure characteristics suitable for the number, load, oil temperature, etc. can be selected. This eliminates the waste of work as a pump and minimizes engine power loss.

  Moreover, in the case of the oil pump 10, the operation control of the cam ring 17 as described above does not require complicated control such as duty control, and the solenoid valve 40 is simply controlled by ON-OFF, and further, the solenoid It can be easily realized with a simple structure using a general solenoid valve 40 without requiring precision processing such as port shape in the valve 40 or tuning of valve opening characteristics. As a result, the manufacturing cost of the pump can be reduced.

  Further, in the oil pump 10, as indicated by a thick solid arrow in FIG. 3, the internal pressure of each pump chamber 20 related to the discharge region acts on the inner peripheral surface of the cam ring 17 on the pivot portion 17a side. The cam ring 17 is pressed in the right direction in the drawing along the cam ring reference line M, that is, toward the support groove 11b, and the pivot portion 17a is pressed against the support groove 11b. However, in the case of the oil pump 10 according to the present embodiment, these pumps are disposed in the outer peripheral region of the cam ring 17 on the pump discharge side, that is, with the peripheral wall of the cam ring 17 interposed between the pump chambers 20 in the discharge region. Since both the pressure chambers 31 and 32 are disposed so as to face the chamber 20, the internal pressures of both the pressure chambers 31 and 32 support the cam ring 17, respectively, as indicated by thick broken line arrows in FIG. 3. It acts to push back to the opposite side to the groove 11b, and the pressure contact of the pivot portion 17a with the support groove 11b is reduced. Thereby, friction between the pivot portion 17b and the support groove 11b when the cam ring 17 is eccentrically swung can be reduced. As a result, it becomes possible to suppress the wear of the pivot portion 17a and the support groove 11b, particularly the wear of the support groove 11b made of a material having a lower hardness than the cam ring 17, thereby improving the durability of the pump.

  Although this action almost cancels out the force acting on the inner and outer peripheral sides of the cam ring 17 on the pump discharge side, the outer peripheral area of the cam ring 17 on the pump suction side located on the opposite side of the support groove 11b Atmospheric pressure or suction pressure acts through the introduction passage 24, and the pivot portion 17a is pressed into the support groove 11b by the atmospheric pressure or suction pressure. Therefore, the pivot portion 17a is supported by the support groove 11b. There is no risk of being separated from the inner surface of the. Accordingly, it is possible to obtain an appropriate operation of the cam ring 17 in which the pivot portion 17a is appropriately slidably contacted with the support groove 11b.

  Further, as described above, since the pressure chambers 31 and 32 are disposed in the region on the pump discharge side so as to face the pump chambers 20 related to the discharge region, the cam ring 17 is disposed in the region. Since the pressure acting on the inner circumferential side and the pressure acting on the outer circumferential side are all substantially equal to the discharge pressure, the pressure difference between the inner and outer circumferences of the cam ring 17 in the ejection area can be minimized. As a result, in the discharge region, leakage of lubricating oil through the minute gap interposed between the both side surfaces of the cam ring 17 and the bottom wall 13a of the pump housing chamber 13 and the inner surface of the cover member 12 is minimized. It becomes possible to suppress to. As a result, the waste of work of the oil pump 10 is reduced as much as possible, and the efficiency of the oil pump 10 can be increased.

  As described above, according to the oil pump 10 according to the present embodiment, the first and second pressure chambers 31 and 32 are arranged on both sides of the pivot portion 17 a, so that the internal pressure of the second pressure chamber 32 is applied to the spring 18. Since it acts so as to assist the force, the biasing force of the spring 18 can be set as small as possible. Specifically, the arrangement of the second pressure chamber 32 allows the spring 18 to have an urging force that can ensure the low-pressure characteristic X, that is, an urging force that is balanced with the first operating oil pressure Px. Therefore, it is possible to use a low-load spring having a smaller spring constant than conventional ones. Thereby, the space required for arrangement | positioning of the spring 18 in the pump body 11 can be reduced, and the said oil pump 10 can be reduced in size and weight. As a result, the mountability of the oil pump 10 to the engine is improved.

  In addition, since the second pressure receiving surface 34 is set to a pressure receiving area smaller than that of the first pressure receiving surface 33, the hydraulic pressure of the cam ring 17 can be set in two stages by the second pressure chamber 32. Become. Thereby, the freedom degree of the discharge pressure characteristic of a pump can also be improved.

  Various pumps configured to control the swing of the cam ring by the differential pressure between the two pressure chambers, such as a variable displacement pump for a power steering device, have been provided. Each has a structure for generating a differential pressure based on a pressure loss caused by an orifice or the like, and this pressure loss reduces pump efficiency. On the other hand, in the case of the oil pump 10 according to the present embodiment, the discharge pressure is introduced into the first pressure chamber 31 and the second pressure chamber 32 without any pressure loss. Since the structure is such that the operating torque of the cam ring 17 is generated by the pressure receiving area difference of 32, that is, the area difference between the first pressure receiving surface 33 and the second pressure receiving surface 34, a pump such as the conventional variable displacement pump is used. There is no risk of reducing efficiency. As a result, compared to the conventional variable displacement pump, the pressure loss does not occur, which can contribute to improvement of pump efficiency.

  Furthermore, in the oil pump 10 according to the present embodiment, since the solenoid valve 40 is set to have the high pressure characteristic when not energized, even if the solenoid valve 40 fails, the engine It also has a fail-safe function that ensures the required discharge pressure over the entire use area.

  10 and 11 show a modification of the first embodiment, and the solenoid valve 40 according to the first embodiment is configured as a so-called normally closed type.

  In other words, the solenoid valve 40 according to this modification is configured as a so-called normally closed type having characteristics opposite to those of the first embodiment, and when not energized as shown in FIG. The IN port 51a is blocked and the OUT port 51b communicates with the drain port 51c. As shown in FIG. 11, the IN port 51a and the OUT port 51b communicate with each other when energized. Thus, the oil pump 10 has the low pressure characteristic X when the solenoid valve 40 is not energized, and the high pressure characteristic Y when the solenoid valve 40 is energized.

  According to this configuration, when the frequency of the high pressure characteristic Y is lower than the frequency of the low pressure characteristic X of the discharge pressure characteristic of the oil pump 10 required for the engine, the energization time for the solenoid valve 40 can be shortened. Therefore, the solenoid valve 40 is used for suppressing deterioration over time.

  12 to 16 show a second embodiment of the oil pump according to the present invention. The arrangement of the seal members 30 and 30 according to the first embodiment is changed, and the solenoid valve 40 is integrated with the housing. It is composed.

  That is, in the present embodiment, the seal holding grooves 17e and 17f provided in the seal constituent portions 17c and 17d of the cam ring 17 in the first embodiment are abolished, and instead the seal sliding contact surfaces 11c. 11d, seal holding grooves 11e, 11f similar to the respective seal holding grooves 17e, 17f are formed at positions facing the abolished seal holding grooves 17e, 17f, and the respective seal holding grooves 11e, In 11f, the respective elastic members 29, 29 and the respective seal members 30, 30 are accommodated and arranged.

  In this embodiment, as shown in FIGS. 12, 15 and 16, the valve body 41 of the solenoid valve 40 is integrally formed on the outer surface 12b of the cover member 12 in parallel with the cam ring eccentric direction line N. The solenoid valve 40 is formed integrally with the housing. The structure of the solenoid valve 40 is the same as that of the first embodiment, and the valve body 41 is slidably accommodated in the valve body 41 formed integrally with the cover member 12. The electromagnetic unit 44 is attached to one end opening which is the upper end in FIG.

  Further, as shown in FIG. 16, the inner surface 12c of the cover member 12 has a suction port 21, a discharge port 22, a communication groove 23 that connects the discharge port 22 and the bearing hole 12a, and An introduction passage 25 extending from the discharge port 22 is provided similarly to the pump body 11.

  Further, the cover member 12 is provided with an IN port 41a at a predetermined position of the introduction passage 24 so as to communicate the inside of the pump body 11 (pump housing chamber 13) and the inside of the valve body 41. OUT ports 41b that also serve as the introduction holes 35 are formed in predetermined positions that are substantially symmetrical with respect to the cam ring reference line M with respect to 41a, and the valve body 11 that is integrally formed with the cover member 12 is provided. A drain port 41c and a back pressure port 41d are formed at predetermined positions on the peripheral wall and the bottom wall, respectively.

  Therefore, according to this embodiment, when the cam ring 17 is eccentrically swung, each of the seal members 30 and 30 is made of an iron-based sintered material having a hardness higher than that of the pump body 11 made of an aluminum alloy material. Therefore, the wear of the mating member by the seal members 30 and 30 can be suppressed. As a result, the durability of the oil pump 10 is improved compared to the first embodiment.

  Furthermore, in the present embodiment, the solenoid valve 40 is formed integrally with the cover member 12, that is, the housing, so that the hydraulic circuit of the oil pump 10 shown in FIG. 8 can be completed within the oil pump 10. For this reason, it serves for miniaturization as a hydraulic pressure supply system centering on the oil pump 10.

  FIGS. 17 and 18 show a third embodiment of the oil pump according to the present invention. Based on the configuration of the second embodiment, each seal component 17c of the cam ring 17 according to the second embodiment. 17d, the respective seal holding grooves 17e, 17f, the respective elastic members 29, 29 and the respective seal members 30, 30 accommodated therein are eliminated.

  That is, in this embodiment, instead of eliminating the respective seal members 30, 30, etc., the side surface 17j of the first seal constituting portion 17c of the cam ring 17 is formed in a flat shape, and the bolt insertion in each figure is made. The seal constituting portion 11h constituting the seal portion SL by contacting the side surface 17j of the first seal constituting portion 17c of the cam ring 17 at the maximum eccentricity of the cam ring 17 on the inner side portion of the pump body 11 in the vicinity of the portion 11g. The first seal component 17c is provided so as to face the side surface 17j.

  This seal component portion 11h is provided so as to be in close contact with the side surface 17j of the first seal component portion 17c of the cam ring 17 when the cam ring 17 is fully eccentric. The seal component portion 11h moves the cam ring 17 in the eccentric direction. The movement is restricted and the inside of the first pressure chamber 31 is liquid-tightly held by the seal portion SL constituted by the seal constituting portion 11h. In addition, with the above configuration change, in the present embodiment, the support protrusion 11i provided in the first embodiment to restrict the maximum eccentric position of the cam ring 17 on the inner side of the pump body 11 is eliminated. .

  According to such a configuration, when the cam ring 17 is not operated (at the time of maximum eccentricity), that is, at the stage of increasing the discharge pressure, the second embodiment using the seal members 30 and 30 by the seal portion SL. As a result, the inside of the first pressure chamber 31 can be kept liquid-tight. As a result, the discharge pressure can be increased with an appropriate time (response) up to the first operating oil pressure Px set as the minimum required oil pressure at the time of low engine speed. As a result, the required hydraulic pressure at the time of engine low rotation, such as the required hydraulic pressure P1 of the valve timing control device, can be secured.

  On the other hand, at the time of operation of the cam ring 17 (at the time of eccentric oscillation), that is, at the stage of suppressing an increase in discharge pressure, the pressure chambers 31 and 32 are respectively connected to the seal sliding contact surfaces 11c and 11d and the seals. It is sealed by the minute clearance C formed between the surfaces 17g and 17h. In this case, although some leakage occurs through the minute clearance C, at this stage, the discharge pressure exceeds the first operating oil pressure Px and the increase in the discharge pressure is suppressed. Therefore, the leak can be tolerated.

  The minute clearance C is an axial gap between the rotor 15 and the cam ring 17 in the oil pump 10 according to the embodiment and the inner side surface 12c of the cover member 12 and the bottom wall 13a of the pump housing chamber 13, or a well-known clearance C. In the trochoid pump, the clearance is set to be approximately the same as the radial clearance between the outer peripheral surface of the rotor and the inner peripheral surface of the housing, and in principle, the leak is set to be within an allowable range.

  Therefore, according to this embodiment, since each of the seal members 30 and 30 is abolished, each of the components of the oil pump 10 such as the seal members 30 and 30 and the elastic members 29 and 29 associated with the seal members 30 and 30 is provided. The number of parts can be reduced, and thereby the number of assembling steps of the oil pump 10 can be reduced. As a result, the manufacturing cost of the oil pump 10 is reduced.

  Moreover, by reducing the number of components of the oil pump 10, it is possible to suppress problems associated with the assembly, such as an assembly failure, and the stability and improvement of the quality of the oil pump 10 are also provided.

  19 to 22 show a fourth embodiment of an oil pump according to the present invention. Based on the configuration of the first embodiment, a discharge pressure is substituted for the solenoid valve 40 according to the first embodiment. The discharge pressure characteristic of the pump is changed with the hydraulic direction switching valve 50 operated by the above.

  That is, in this embodiment, a known spool-type hydraulic direction switching valve 50 is used as an alternative to the solenoid valve 40, and the hydraulic direction switching valve 50 has one end as shown in FIGS. A cylindrical valve body 51 having an opening on the other side and closed on the other end side, a plug 52 for closing one end opening of the valve body, and a valve body 51 slidably accommodated in the axial direction. The valve body 53 defining the pressure chamber 55 and the back pressure chamber 56 in the valve body 51 and the back pressure chamber 56 are accommodated in the valve body 51 by the first and second land portions 53a and 53b included in the valve body 51. When the internal pressure of the pressure chamber 54 exceeds a predetermined set pressure Pz set higher than the required hydraulic pressure P1 and lower than the required hydraulic pressure P2, as shown in FIG. , The valve body 53 is set to move to the back pressure chamber 56 side against the biasing force of the spring 54.

  The valve body 51 includes an IN port 51a connected to the discharge hole 22a, an OUT port 51b connected to the introduction hole 35, and a suction port 21 or a drain connected to the outside at predetermined positions in the axial direction on the peripheral wall. Each of the ports 51c is formed so as to penetrate therethrough, and a back pressure port 51d that is connected to the suction port 21 or the outside and constantly opens the back pressure chamber 45 to the suction pressure or the atmospheric pressure passes through the side wall on the back pressure chamber 55 side. Is formed.

  The plug 52 is screwed into a female thread portion provided on an inner peripheral surface of an opening end portion on one end side of the valve body 51, and an introduction port 52a is formed through the shaft center. The introduction port 52a Through this, the discharge pressure is constantly introduced into the pressure chamber 55.

  The valve body 53 is formed with a reduced diameter at an intermediate portion in the axial direction, and an annular space 57 is defined between the land portions 53 a and 53 b and the valve body 51. The port 51b communicates with the IN port 51a or the drain port 51c. That is, when the valve body 53 is in an inoperative state, the IN port 51a is blocked by the first land portion 53a, and the OUT port 51b and the drain port 51c communicate with each other via the annular space 57. When the body 53 is activated, the drain port 51c is blocked by the second land portion 53b, and the IN port 51a and the OUT port 51b communicate with each other through the annular space 57.

  With the configuration as described above, in the case of the oil pump 10 according to the present embodiment, the IN port 51a of the hydraulic direction switching valve 50 is shut off only in the first pressure chamber 31 when the engine speed is low. Since the discharge pressure acts, as shown in FIG. 22, when the discharge pressure reaches the first operating oil pressure Px, the cam ring 17 swings in the direction of decreasing the eccentric amount, and the discharge pressure rises gradually. The low pressure characteristic X is exhibited. (T1 section in the figure). When the discharge pressure rises and the internal pressure of the pressure chamber 55 reaches the set pressure Pz, the valve body 53 is pivoted toward the back pressure chamber 55 against the biasing force of the spring 53 based on the internal pressure of the pressure chamber 55. As the valve body 52 moves in the axial direction, the drain port 51 c is closed by the second land portion 53 b and the IN port 51 a gradually opens into the annular space 57. As a result, the IN port 51a and the OUT port 51b gradually communicate with each other through the annular space 57, and the discharge pressure is gradually introduced into the second pressure chamber 32. As a result, the internal pressure of the second pressure chamber 32 rises, and the cam ring 17 swings in the direction of increasing the eccentric amount accordingly, so that the high pressure characteristic Y such that the discharge pressure further increases is exhibited ( T2 section in FIG. 22).

  Thus, according to this embodiment, unlike the solenoid valve 40 according to the first embodiment, the discharge pressure characteristics cannot be freely switched according to the operating state of the engine, but the engine can be manufactured at a lower manufacturing cost. An oil pump having a discharge pressure characteristic matched to the number of rotations can be obtained.

  The present invention is not limited to the configuration of each of the above embodiments. For example, the required hydraulic pressures P1 to P5, the first and second operating hydraulic pressures Px and Py, and the set pressure Pz are mounted on the oil pump 10. It can be freely changed according to the specifications of the internal combustion engine and valve timing control device of the vehicle to be used.

  Further, in each of the above embodiments, each side wall according to the present invention has been described as an example in which the side wall is configured by the bottom wall of the pump body 11 and the cover member 12, but each side wall may be provided on both sides of the pump element, for example. It is also possible to have a separate structure from the housing, such as being configured as a separate member interposed between the bottom wall of the pump body 11 and the axial direction of the cover member 12.

  In each of the above embodiments, the operation of the cam ring 17 is controlled by balancing the urging force of the spring 18 and the internal pressure of the second pressure chamber 32 with respect to the internal pressure of the first pressure chamber 31. In some cases, the spring 18 is abolished by setting the pressure receiving area of the first pressure receiving surface 33 to be larger than the pressure receiving area of the second pressure receiving surface 34, and the cam ring is formed only by the internal pressure (differential pressure) of the pressure chambers 31 and 32. The operation of 17 may be controlled.

  Further, in each of the above embodiments, the pressure receiving area of the second pressure receiving surface 34 is set smaller than the pressure receiving area of the first pressure receiving surface 33. However, depending on the requirements of the internal combustion engine or the like, the pressure receiving surfaces 33, 34 are both. May be set equal.

11 ... Pump body (housing)
12 ... Cover member (housing)
13 ... Pump housing 14 ... Drive shaft 15 ... Rotor 16 ... Vane 17 ... Cam ring 18 ... Spring (biasing member)
20 ... Pump chamber (hydraulic chamber)
21 ... Suction port (suction part)
22: Discharge port (discharge part)
31 ... 1st pressure chamber 32 ... 2nd pressure chamber 33 ... 1st pressure receiving surface 34 ... 2nd pressure receiving surface 40 ... Solenoid valve (control means)

Claims (1)

A variable displacement oil pump that supplies hydraulic oil to at least each sliding portion of the internal combustion engine,
A pump element composed of a rotor that is rotationally driven by an internal combustion engine and a plurality of vanes provided in an outer periphery of the rotor so as to be able to appear and retract;
A cam ring in which the eccentric amount with respect to the rotation center of the rotor changes by accommodating the pump element on the inner peripheral side and swinging around a swing fulcrum provided on the outer peripheral side;
A plurality of hydraulic oil chambers defined by the rotor and adjacent vanes by arranging side walls on both axial sides of the cam ring;
The cam ring is housed inside, and a discharge portion that opens through the side wall to a discharge region where the volumes of the plurality of hydraulic oil chambers decrease along the rotation direction of the rotor and the rotation direction of the rotor A housing having a suction portion that opens through the side wall in a suction region in which the volumes of the plurality of hydraulic oil chambers increase;
A biasing member that biases the cam ring in a direction that increases the amount of eccentricity of the cam ring with respect to the rotation center of the rotor;
The discharge pressure is defined by the outer peripheral side of the cam ring, the inner peripheral surface of the housing, and the swinging fulcrum of the cam ring, and the discharge pressure is introduced into the inside so as to hinder the biasing force of the biasing member. A first pressure chamber for applying a swinging force in a direction to reduce the amount of eccentricity with respect to the cam ring via a first pressure receiving surface to be applied;
The discharge pressure is defined by the outer peripheral side of the cam ring, the inner peripheral surface of the housing, and the swinging fulcrum of the cam ring, and the discharge pressure is introduced inside to assist the biasing force of the biasing member. A second pressure chamber that imparts a swinging force in a direction that increases the amount of eccentricity with respect to the cam ring via a second pressure receiving surface that acts on the cam ring;
Control means for switching and controlling the supply of the discharge pressure to the second pressure chamber,
The area of the first pressure receiving surface is set larger than the area of the second pressure receiving surface,
The first pressure chamber and the second pressure chamber are provided on the swing fulcrum side with respect to the center of the cam ring so that a part of the chamber overlaps with the discharge region at least in the rotor radial direction. Features variable displacement oil pump.

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