JP5564450B2 - Oil pump - Google Patents

Description

  The present invention relates to an oil pump applied to a hydraulic power source that supplies hydraulic oil to each sliding portion of an internal combustion engine for an automobile, for example.

  As a conventional oil pump applied to an internal combustion engine of an automobile, for example, one described in Patent Document 1 below is known.

  Briefly speaking, this oil pump is a so-called variable displacement vane pump, and the amount of eccentricity of the cam ring urged by a spring in a direction that is always eccentric with respect to the rotation center of the rotor is determined between the housing and the cam ring. Control is performed based on the discharge pressure introduced into the defined control oil chamber, thereby making the discharge amount variable, thereby reducing the drive torque of the pump and saving energy.

JP 2008-309049 A

  Incidentally, in recent years, it has been desired to increase the discharge amount and to reduce the size of the pump by driving the conventional oil pump at a high speed that is higher than the internal combustion engine.

  However, when the conventional oil pump is driven at the above-described high speed, the suction amount cannot catch up and cavitation occurs, which causes a problem that a sufficient discharge amount cannot be secured.

  Accordingly, the present invention has been devised in view of the technical problem of the conventional oil pump described above, and an object thereof is to provide an oil pump that can secure a sufficient discharge amount by suppressing cavitation even during high-speed rotation. It is said.

  In the present invention, in particular, when the pressure in the hydraulic oil chamber that opens to the upstream side of the region in which the internal volume is reduced is a negative rotation speed, the hydraulic oil is supplied from the low pressure portion to the hydraulic oil chamber that is negative pressure. It is characterized by controlling to flow in.

  According to the present invention, when the hydraulic oil chamber that opens to the upstream side of the region where the internal volume is reduced is rotated at a high speed at which the negative pressure is applied, the hydraulic oil is supplied from the low pressure portion to the hydraulic oil chamber that has become the negative pressure. As a result, the suction amount can be increased, so that the occurrence of cavitation is suppressed even during the high-speed rotation, and a sufficient discharge amount can be secured.

It is a front view of the balancer device which constituted the oil pump concerning the present invention integrally. It is sectional drawing which follows the AA line of FIG. It is a top view of the balancer apparatus shown in FIG. It is a bottom view of the balancer apparatus shown in FIG. It is a disassembled perspective view which shows the structure of the oil pump which concerns on 1st Embodiment of this invention. It is a longitudinal cross-sectional view of the oil pump shown in FIG. It is sectional drawing which follows the BB line of FIG. It is the figure which looked at the pump body simple substance shown in FIG. 5 from the mating surface side with a cover member. It is the figure which looked at the cover member single-piece | unit shown in FIG. 5 from the mating surface side with a pump body. It is the graph showing the relationship between the spring load of both springs shown in FIG. 7, and a cam ring rocking | fluctuation angle. It is a rear view of the oil pump shown in FIG. It is sectional drawing which follows the CC line of FIG. It is the figure which looked at the control valve (valve body) shown in FIG. 12 from the mating surface side with a cover member. It is a hydraulic circuit diagram of the oil pump concerning the embodiment, (a) shows the 1st state concerning the present invention, and (b) shows the 2nd state concerning the present invention. It is a graph showing the hydraulic characteristic of the oil pump which concerns on the same embodiment. 15 is a diagram showing the phase of the cam ring and the state of the hydraulic circuit in each section shown in FIG. 15, where (a) is a section a, (b) is a section b, (c) is a section c, and (d) is a section d. , (E) show the section e, respectively. It is a hydraulic-circuit figure of the oil pump which concerns on 2nd Embodiment of this invention, Comprising: (a) has shown the 1st state which concerns on this invention, (b) has shown the 2nd state which concerns on this invention. . It is a hydraulic circuit diagram of the oil pump concerning a 3rd embodiment of the present invention, and (a) shows the 1st state concerning the present invention, and (b) shows the 2nd state concerning the present invention. .

  Below, each embodiment of the oil pump concerning the present invention is explained in full detail based on a drawing. In each of the following embodiments, this oil pump is applied as an oil pump that supplies lubricating oil for an engine to a valve timing control device that controls the opening and closing timing of a sliding portion of an automobile internal combustion engine and an engine valve. Show.

  1 to 16 show a first embodiment of an oil pump according to the present invention. This oil pump 10 is a balancer device provided at the lower part of the internal combustion engine in order to reduce secondary vibration of the internal combustion engine. 1 and is driven by the balancer device 1.

  Therefore, before explaining the oil pump 10, first, the balancer device 1 will be briefly described. The balancer device 1 is not shown in FIG. An upper housing 2 that is integrally formed with the oil pan, a lower housing 3 that is attached to the lower portion of the upper housing 2, and the two housings 2, 3 A drive-side drive shaft 4 which is a pair of balance shafts arranged in parallel along the axial direction of the crankshaft 6 which is rotatably supported between them and which is the longitudinal direction of the engine, and a driven-side driven shaft which is driven by the balance shaft 5 and the axial ends of the shafts 4 and 5 (the right side end in FIG. 2) are fixed to each other and meshed with each other. The drive shaft 4 is mainly composed of a rotating helical drive gear 4 a and a driven gear 5 a, and the drive shaft 4 is driven at a rotational speed twice that of the crankshaft 6 with the power of the crankshaft 6 transmitted through the chain 7. On the other hand, the driven shaft 5 is driven at the same rotational speed as the drive shaft 4 in the opposite direction to the drive shaft 4 via the both gears 4a and 5a.

  The oil pump 10 is attached to the front end portion of the balancer device 1, and its drive shaft 14 is connected to the other axial end portion of the driven shaft 5, and is rotationally driven with power transmitted from the driven shaft 5. The Rukoto. The lower housing 3 is provided with an oil strainer 8 for sucking in the hydraulic oil stored in the oil pan. The hydraulic oil sucked up by the oil strainer 8 is provided in the lower housing 3. The oil pump 10 is led to a later-described suction port 21a through a suction passage (not shown). On the other hand, the upper housing 2 is provided with a discharge hole 2a for discharging (introducing) hydraulic oil to a cylinder block (not shown), and the hydraulic oil discharged from the discharge port 22a described later of the oil pump 10 is It is guided to the discharge hole 2a through a discharge passage (not shown) formed inside the housings 2 and 3, and is discharged to the cylinder block through the discharge hole 2a.

  As described above, since the oil pump 10 is configured integrally with the balancer device 1, hydraulic oil discharged from the oil pump 10 can be supplied to the internal combustion engine through the balancer device 1. There is no need to provide piping or the like for connecting the oil pump 10 and the internal combustion engine, and the structure around the engine is simplified.

  Next, the oil pump 10 according to the present invention will be described in detail. As shown in FIGS. 5 to 7, the oil pump 10 is formed with an opening at one end side and accommodates a pump structure described later therein. A pump housing 11 (port block according to the present invention) comprising a pump body 11 having a substantially U-shaped longitudinal section and a cover member 12 that closes one end opening of the pump body 11. A drive shaft 14 that is rotatably supported and is driven to rotate by the balancer device 1 through substantially the center of the pump housing chamber 13, and is housed in the pump housing chamber 13. Is driven in a counterclockwise direction to increase or decrease the volume of the pump chamber PR, which is a plurality of hydraulic oil chambers formed therein. It includes a pump structure that performs use, the.

  Here, the pump structure is rotatably accommodated in the pump accommodating chamber 13, and a rotor 15 whose central portion is coupled to the outer periphery of the drive shaft 14 and a radially notched outer periphery of the rotor 15 are formed. And a vane 16 accommodated in the plurality of slits 15a, respectively, and arranged on the outer peripheral side of the rotor 15 so as to be eccentric with respect to the rotation center of the rotor 15, and the pump housing, the rotor 15 and two adjacent two A cam ring 17 that defines the pump chambers PR together with the vanes 16 and 16, and a pair of ring members 18 that are formed to have a smaller diameter than the rotor 15 and are slidably disposed on both inner peripheral sides of the rotor 15. , 18.

  The pump body 11 is integrally formed of an aluminum alloy material. As shown in FIGS. 5 to 8, the pump body 11 is substantially at the center position of the end wall 11 a of the pump body 11 that forms one end wall of the pump housing chamber 13. A bearing hole 11b that rotatably supports one end of the drive shaft 14 is formed. Further, a support groove 11c having a substantially semicircular cross section for supporting the cam ring 17 through a pivot pin 19 so as to be swingable is formed at a predetermined position on the inner peripheral wall of the pump chamber 13. Further, on the inner peripheral wall of the pump housing chamber 13, on the Y axis negative direction side in FIG. 7 with respect to a straight line (hereinafter referred to as “cam ring reference line”) M connecting the center of the bearing hole 11 b and the center of the support groove 11 b. A seal slidable contact surface 11 d is formed on the outer peripheral portion of the cam ring 17 to which a later-described seal member 20 is slidably contacted. The seal slidable contact surface 11d is formed in a circular arc shape having a predetermined radius R1 from the center of the support groove 11c, and the seal member 20 can always slidably contact within a range in which the cam ring 17 swings eccentrically. When the cam ring 17 is oscillated eccentrically, the cam ring 17 is slidably guided along the seal sliding contact surface 11d, so that the cam ring 17 can be smoothly operated (eccentric oscillation). It is supposed to be.

  Further, on the inner side surface of the end wall 11a of the pump body 11, particularly as shown in FIGS. 7 and 8, in the outer peripheral region of the bearing hole 11b, the pump chamber PR is provided with the pump action by the pump structure. A substantially arc-shaped suction port 21 into which hydraulic oil is introduced from the balancer device 1 side through a suction port 21a, which will be described later, so as to open to a region where the volume is expanded (hereinafter referred to as "volume expansion region") is formed as a notch. Has been. On the other hand, in the area where the volume of each pump chamber PR is reduced (hereinafter referred to as “volume reduction area”), the hydraulic oil discharged from each pump chamber PR is supplied downstream to the discharge port 22a. A substantially arc concave discharge port 22 for guiding the hydraulic oil has a substantially arc concave shape that can be switched to function as the intake port 21 or the discharge port 22 according to the operating hydraulic pressure in the engine, that is, the discharge pressure, upstream of the discharge port 22. The switching ports 23 are notched so as to face the suction port 21 with the bearing hole 11b interposed therebetween.

  The suction port 21 is integrally provided with an introduction portion 24 formed so as to bulge toward the first spring accommodating chamber 28 described later at a substantially intermediate position in the circumferential direction. 21 is connected to the suction passage in the balancer device 1 by passing through the end wall 11a of the pump body 11 and opening to the outside at a position near the boundary with the suction port 21 and on the start end side of the suction port 21. The suction port 21a is formed through. As a result, the hydraulic oil sucked up from the oil pan through the oil strainer 8 and guided to the suction port 21a through the suction passage is introduced into the introduction portion 24 and the suction portion by the negative pressure generated by the pump action by the pump structure. It will be supplied to each pump chamber PR located in the volume expansion region via the port 21. The suction port 21a is configured to communicate with a low pressure chamber 36, which will be described later, formed in the outer peripheral region of the cam ring 17 together with the introduction portion 24, and the low pressure chamber 36 also operates at a low pressure that is the suction pressure. It comes to guide the oil. Here, the low-pressure part according to the present invention is a concept including the entire range of the suction pressure communicated with the suction port 21, but mainly the suction port 21 and the introduction parts 24 to 24 adjacent thereto. The low pressure chamber 36 is configured.

  The discharge port 22 has a discharge port 22a penetratingly formed on the terminal side thereof, penetrating through the end wall 11a of the pump body 11 and opening to the outside so as to be connected to the discharge passage in the balancer device 1. . As a result, hydraulic oil that is pressurized by the pump action of the pump structure and discharged to the discharge port 22 is provided in the cylinder block from the discharge port 22a through the discharge passage and the like in the cylinder block. It will be supplied to each sliding part in an engine, a valve timing control device, etc. via this. Here, the discharge port 22a is provided so as to bulge radially outward with respect to both ends in the circumferential direction of the discharge port 22, and the outer half side thereof is formed in the cam ring 17 as will be described later. It communicates with a later-described discharge port 32 (second communication hole 32a) provided on the cover member 12 side.

  Further, a communication groove 25 that connects the discharge port 22 and the bearing hole 11b is formed in the vicinity of the start end of the discharge port 22, and hydraulic fluid is supplied to the bearing hole 11b via the communication groove 25. By supplying hydraulic oil to the sides of the rotor 15 and the vanes 16 as well as supplying them, good lubrication of each sliding portion is ensured. Note that the communication groove 25 is formed so as not to coincide with the direction in which each vane 16 protrudes and is prevented from dropping into the communication groove 25 when the vanes 16 appear and disappear. Furthermore, one end of an internal passage 26 that is provided inside the end wall 11 a of the pump body 11 and communicates with the discharge port 22 and a control oil chamber 37 described later is formed at the start end of the discharge port 22. A part of the hydraulic oil discharged to the discharge port 22 is guided to the control oil chamber 37 described later through the internal passage 26.

  The switching port 23 is integrally provided with a communication portion 27 formed so as to bulge outward in the radial direction at the terminal end position, and is formed inside the cam ring 17 via the communication portion 27. Through the second communication passage 39 described later, the switch port 33 (third communication hole 33a) on the cover member 12 side is communicated.

  As shown in FIGS. 5 and 9, the cover member 12 has a substantially plate shape and is attached to the opening end surface of the pump body 11 by a plurality of bolts B <b> 1, and faces the bearing hole 11 b of the pump body 11. A bearing hole 12a that rotatably supports the other end side of the drive shaft 14 is formed through the position. In addition, the inner surface of the cover member 12 that constitutes the other end wall of the pump housing chamber 13 is also located at a position facing the ports 21 to 23 provided in the pump body 11, and substantially the ports 21 to 23. A suction port 31, a discharge port 32, and a switching port 33 that are similarly configured are each formed with a notch. A first communication hole 31 a connected to the first port 46 of the control valve 40 described later is provided at the end portion of the suction port 31, and a second communication valve 31 described later is provided in the middle portion of the discharge port 32. A second communication hole 32 a connected to the port 47 is formed, and a third communication hole 33 a connected to a third port 48 of the control valve 40 described later is formed through the terminal portion of the switching port 33. .

  As shown in FIG. 2, the drive shaft 14 has an axial one end 14 a that penetrates the end wall 11 a of the pump body 11 and faces the outside, and rotates integrally with the other axial end of the driven shaft 5 of the balancer device 1. The rotor 15 is connected in a possible manner, and the rotor 15 is rotated counterclockwise in FIG. 6 based on the rotational force transmitted from the driven shaft 5. Here, as shown in FIG. 7, a straight line (hereinafter referred to as “cam ring eccentric direction line”) N passing through the center of the drive shaft 14 and orthogonal to the straight line M is defined as a volume expansion region and a volume reduction region. When the control valve 40, which will be described later in the normal state, is in the first state, the X-axis negative direction side of the straight line N is the suction region, and the positive direction side is the discharge region.

  As shown in FIGS. 5 to 7, the rotor 15 has a plurality of slits 15 a that are radially formed radially outward from the center side, and an inner base end portion of each slit 15 a. Are provided with back pressure chambers 15b each having a substantially circular cross section for introducing the discharge oil, and the vanes 16 are caused by the centrifugal force accompanying the rotation of the rotor 15 and the pressure in the back pressure chamber 15b. It is pushed out to the outside.

  Each vane 16 is configured such that, when the rotor 15 rotates, each distal end surface is in sliding contact with the inner peripheral surface of the cam ring 17 and each proximal end surface is in sliding contact with the outer peripheral surface of each of the ring members 18 and 18. Yes. That is, each of the vanes 16 is configured to be pushed up radially outward of the rotor 15 by the ring members 18 and 18, the engine speed is low, and the centrifugal force and the pressure of the back pressure chamber 15b are set. Is small, each tip is in sliding contact with the inner peripheral surface of the cam ring 17 so that the pump chambers PR are liquid-tightly separated.

  The cam ring 17 is integrally formed of a so-called sintered metal in a substantially cylindrical shape, and a pivot portion having a substantially arcuate groove shape that forms an eccentric rocking fulcrum by fitting with a pivot pin 19 at a predetermined position on the outer periphery thereof. 17a is cut out along the axial direction, and an arm portion 17b linked to each of the springs 34 and 35, which will be described later, is provided at a position opposite to the pivot portion 17a across the center of the cam ring 17 in the radial direction. It protrudes along. On both sides of the arm portion 17b in the rotation direction, there are a first spring 34 set to a predetermined spring constant and a second spring 35 set to a spring constant smaller than the first spring 34. Opposed. Here, the arm portion 17b is provided with a substantially arc-shaped pressing protrusion 17c on one side in the rotational direction, while the thickness of a restricting portion 30 described later is provided on the other side. The pressing protrusion 17d set longer than the width is extended, and the pressing protrusion 17c is always in contact with the tip of the first spring 34, and the pressing protrusion 17d is always in contact with the tip of the second spring 35. Thus, the arm portion 17b and the springs 34 and 35 are linked to each other.

  Further, as shown in FIGS. 7 and 8, the first and second springs 28 and 29 are accommodated and held in the pump body 11 at positions facing the support groove 11b. The second spring accommodating chambers 28 and 29 are provided adjacent to the pump accommodating chamber 13 along the Y-axis direction in FIG. 7, and the first spring accommodating chamber 28 has an end wall and an arm portion 17b ( The first spring 34 is elastically mounted with a predetermined set load W1 between the pressing protrusion 17c) and the second spring accommodating chamber 29 has an end wall and an arm portion 17b (pressing protrusion 17d). In the meantime, the second spring 35 set to have a smaller wire diameter than the first spring 34 is elastically mounted with a predetermined set load W2. Between the first and second spring accommodating chambers 28 and 29, a restricting portion 30 configured to have a stepped diameter is provided, and the arm portion 17b is provided on one side of the restricting portion 30. When the side portion abuts, the rotation range of the arm portion 17b in the counterclockwise direction is restricted, while the tip of the second spring 35 abuts on the other side portion of the restriction portion 30 so that the second portion. The maximum extension amount of the spring 35 is regulated.

  Thus, with respect to the cam ring 17, the arm portion 17b is moved with the resultant force W0 of the set loads W1, W2 of the springs 34, 35, that is, the biasing force of the first spring 34 that exerts a relatively large spring load. As shown in FIG. 7, the pressing protrusion 17d of the arm portion 17b is in the second state by being constantly biased in the direction in which the amount of eccentricity increases (counterclockwise in FIG. 7). The second spring 35 is compressed by entering the spring accommodating chamber 29, and the other side portion of the arm portion 17b is pressed against one side portion of the restricting portion 30, thereby maximizing the amount of eccentricity. Regulated in position.

  Further, as shown in FIG. 7, the outer periphery of the cam ring 17 has a seal surface 17f concentric with the seal slide contact surface 11d formed to face the seal slide contact surface 11d of the pump body 11. A seal constituting portion 17e having a substantially triangular cross section is projected, and a seal holding groove 17g having a substantially rectangular cross section is formed in a cutout along the axial direction on the seal surface 17f of the seal constituting portion 17e. The seal holding groove 17g accommodates and holds the seal member 20 that is in sliding contact with the seal sliding contact surface 11d when the cam ring 17 is eccentrically swung. Here, the seal surface 17f is constituted by a predetermined radius R2 slightly smaller than the radius R1 constituting the seal sliding contact surface 11d, and between the seal sliding contact surface 11d and the seal surface 17f, A predetermined minute clearance is formed.

  The seal member 20 is formed of, for example, a long and thin linear shape along the axial direction of the cam ring 17 with a fluorine-based resin material having a low friction characteristic, and is disposed at the bottom of the seal holding groove 17g. By being pressed against the seal sliding contact surface 11d by the elastic force of 20a, the seal sliding contact surface 11d and the seal surface 17f are in a state of being liquid-tightly separated.

  Further, both the spring accommodating chambers 28 and 29 are configured to communicate with the low pressure chamber 36 particularly through the introduction portion 24 and the suction port 21, and the cam ring 17 is operated by the hydraulic pressure in the low pressure chamber 36. By being always pressed against the pivot pin 19, the cam ring 17 is eccentrically swung by the pivot pin 19 and the seal member 20 in the outer peripheral region of the cam ring 17 corresponding to the outer peripheral region of the switching port 23. A control oil chamber 37 is provided for control.

  The hydraulic oil in the discharge port 22 is always introduced into the control oil chamber 37 through the internal passage 26, and the pressure receiving surface formed by the side surface of the seal component 17 e facing the control oil chamber 37. By applying a pump discharge pressure to 17h, a swinging force (moving force) is applied to the cam ring 17 in a direction (clockwise in FIG. 7) in which the amount of eccentricity is reduced. That is, the control oil chamber 37 moves the cam ring 17 in the concentric direction by constantly biasing the cam ring 17 in the direction in which the center approaches the concentricity with the rotation center of the rotor 15 via the pressure receiving surface 17h. It is used for quantity control.

  With the configuration as described above, in the oil pump 10, the biasing force in the eccentric direction based on the spring load of the first spring 34 and the biasing force in the concentric direction based on the spring load of the second spring 35 and the internal pressure of the control oil chamber 37. Are balanced with a predetermined force relationship, and the resultant force of the set loads of the two springs 34, 35 is the difference between the set load W1 of the first spring 34 and the set load W2 of the second spring 35. When the urging force based on the internal pressure of the control oil chamber 37 is small with respect to W0 (= W1-W2), the cam ring 17 is in the maximum eccentric state as shown in FIG. When the urging force based on the internal pressure exceeds the resultant force W0 of the set load of the springs 34 and 35, the cam ring 17 is concentric in accordance with the discharge pressure. So that the moving.

  Here, the relationship between the spring load W of both the springs 34 and 35 and the swing angle (movement amount) X of the cam ring 17 will be described in detail. As shown in FIG. When the urging force based on the internal pressure of the control oil chamber 37 reaches the resultant force W0 of the set loads W1, W2 of the springs 34, 35 corresponding to the urging force based on the first operating oil pressure Pf described later at the position X1, As the spring 34 contracts, the second spring 35 begins to expand, and the cam ring 17 moves in the concentric direction. Eventually, when the urging force based on the internal pressure of the control oil chamber 37 increases due to an increase in the pump discharge pressure and the second spring 35 comes into contact with the restricting portion 30, the assisting action by the second spring 35 is lost. Stops moving in the concentric direction (position X2 in the figure). When the discharge pressure further increases and the urging force based on the internal pressure of the control oil chamber 37 reaches the spring load Wx of the first spring 34 corresponding to the urging force based on the second operating oil pressure Ps described later, the first spring 34 is reached. Starts to contract further, and the cam ring 17 further moves in the concentric direction (position X3 in the figure).

  Further, as shown in FIGS. 5 and 11 to 13, the oil pump 10 is controlled by a discharge pressure introduced through the first communication passage 38 provided in the cam ring 17 at the back of the cover member 12. A control valve 40 (corresponding to the control means according to the present invention) is provided for switching and controlling the function of the switching port 23 in accordance with the axial position of the valve body 43.

  In particular, as shown in FIG. 12, the control valve 40 includes a valve body 41 that is formed in a substantially cylindrical shape that is open at one end and closed at the other end, and a plug 42 that closes one end opening of the valve body 41. And a valve having a first land portion 43a and a second land portion 43b slidably in contact with the inner peripheral surface of the valve body 41 at each end thereof. The valve body 43 is elastically mounted with a predetermined set load Wk equal to a port switching hydraulic pressure Pk described later between the plug 42 and the valve body 43 on the inner periphery of one end side of the valve body 41. And a valve spring 44 that is constantly urged toward the other end side.

  The valve body 41 is set to have substantially the same diameter as each of the land portions 43a and 43b of the valve body 43, and the valve body housing portion 41a that houses the valve body 43 and the other end portion via the step portion 41c. A back pressure chamber constituting portion 41b that is provided with a stepped diameter with respect to the valve body housing portion 41a and is separated by a second land portion 43b of the valve body 43 to constitute a back pressure chamber 45 therein. The cover member 12 is fixed to the back surface by a plurality of bolts B2. The peripheral wall of the valve body accommodating portion 41a has a first port 46 connected to the first communication hole 31a via a suction pressure introduction groove 46a provided on the inner surface of the valve body 41, and an inner surface of the valve body 41. A second port 47 connected to the second communication hole 32a via a discharge pressure introduction groove 47a provided in the second communication port 33 and a third port connected to the third communication hole 33a by opening directly to the third communication hole 33a. Each port 48 is formed through. Further, the peripheral wall of the back pressure chamber constituting portion 41b is always connected to the second port 47 via a communication groove 49a provided on the inner surface of the valve body 41, and the discharge ports 22 are connected through the second communication holes 32a. The back pressure port 49 (discharge pressure introduction port according to the present invention) through which the hydraulic oil (discharge pressure) discharged to the gas is always introduced is formed through.

  The valve body 43 is formed with a reduced diameter at an intermediate portion in the axial direction, and an annular space 50 is separated from the valve body 41 by both the land portions 43a and 43b. The port 48 communicates with the first port 46 or the second port 47.

  14A, the control valve 40 has a low discharge pressure introduced into the back pressure chamber 45, and the urging force based on the internal pressure of the back pressure chamber 45 is a valve spring. In a state smaller than the set load Wk of 44, the valve body 43 (second land portion 43 b) is pressed against the stepped portion 41 c of the valve body 41 with the urging force of the valve spring 44. As a result, the first port 46 is blocked by the first land portion 43a, and the second port 47 and the third port 48 communicate with each other via the annular space 50, whereby the switching ports 23, 33 function as discharge ports. (First state according to the present invention).

  On the other hand, as shown in FIG. 14 (b), the discharge pressure introduced into the back pressure chamber 45 increases as the rotational speed of the internal combustion engine, that is, the rotational speed of the oil pump 10 increases, and the internal pressure of the back pressure chamber 45 increases. When the urging force based on the valve spring 44 becomes larger than the set load Wk of the valve spring 44, the valve body 43 resists the urging force of the valve spring 44 with the urging force based on the discharge pressure (the plug 42). Side). As a result, the first port 46 communicates with the third port 48 via the annular space 50, and the second port 47 communicates with the back pressure port 49, whereby the switching ports 23 and 33 function as suction ports. (The second state according to the present invention).

  Here, in the port switching control of the control valve 40, the pressure receiving area of the valve body 43 (second land portion 43b) and the set load Wk of the valve spring 44 are such that the internal pressures of the switching ports 23 and 33 are negative. The urging force of the back pressure chamber 45 based on the discharge pressure immediately before the urging force and the urging force of the valve spring 44 are set so as to be substantially balanced. In the first state, the switching ports 23 and 33 When the internal pressure is negative, the configuration has already been switched from the first state to the second state.

  Hereinafter, the characteristic operation of the oil pump 10 according to the present embodiment will be described with reference to FIGS. 15 and 16.

  First, before entering into the explanation of the operation of the oil pump 10, the required oil pressure of the internal combustion engine, which serves as a reference for the discharge pressure control of the oil pump 10, will be described with reference to FIG. 15. In FIG. P2 in the figure is the required hydraulic pressure of the device when the oil jet used for cooling the piston is employed. , P3 in the figure indicates the third engine required oil pressure required for the lubrication of the bearing portion of the crankshaft at the time of high engine rotation, and is a diagram connecting these points P1 to P3. A middle curve E (two-point difference line) represents an ideal required oil pressure (discharge pressure) P corresponding to the engine speed R of the internal combustion engine. Further, Pf in the figure is the first hydraulic pressure at which the cam ring 17 starts to swing against the resultant force W0 of the springs 34 and 35 by the urging force based on the internal pressure of the control oil chamber 37. Ps indicates the second hydraulic pressure at which the cam ring 17 starts to further swing against the spring load W1 of the first spring by the urging force based on the internal pressure of the control oil chamber 37, respectively. Pk set between the operating hydraulic pressures Pf and Ps indicates a port switching hydraulic pressure for switching the connection between the switching ports 23 and 33.

  That is, in the case of the oil pump 10, in the section a in FIG. 15 corresponding to the idling operation of the engine, the urging force by the resultant force W0 of both the springs 34, 35, that is, the relatively large spring load of the first spring 34 is obtained. As shown in FIG. 16A, the biasing force based on the cam ring 17 keeps the eccentric amount of the cam ring 17 at a maximum. Accordingly, in the section a, the discharge pressure (in-engine hydraulic pressure) P is sufficiently smaller than the port switching hydraulic pressure Pk, and the valve body 43 of the control valve 40 is located closest to the back pressure chamber 45 side. When the switching ports 23 and 33 are connected to the discharge ports 22 and 32, respectively, the discharge pressure P has a characteristic of increasing so as to be substantially proportional to the engine speed R.

  Thereafter, when the engine speed R rises and the discharge pressure P reaches the first operating oil pressure Pf set higher than the first engine required oil pressure P1, as shown in FIG. Due to the urging force based on the internal pressure and the urging force of the second spring 35, the cam ring 17 starts to move in the concentric direction against the urging force of the first spring 34. Then, the increase in the discharge amount is limited by gradually decreasing the eccentric amount of the cam ring 17, thereby suppressing the increase in the discharge pressure P based on the increase in the engine speed R (section in FIG. 15). b). Eventually, as the cam ring 17 moves, the second spring 35 expands, and as shown in FIG. 16C, when the tip of the second spring 35 comes into contact with the regulating portion 30, the assisting action by the second spring 35 is lost. The movement of the cam ring 17 is stopped, and the discharge pressure P increases again in proportion to the increase in the engine speed R (section c in FIG. 15).

  Here, in the present embodiment, the oil pump 10 is not the rotational speed of the internal combustion engine (crankshaft) as in the prior art, but the rotational speed of the balancer device 1 (driven shaft 5), that is, the rotational speed that is twice the conventional rotational speed. Therefore, in the oil pump 10, cavitation is inevitably generated in the region of the engine speed R exceeding the predetermined speed Rk, and thereby the discharge pressure P is increased by the engine speed R. It becomes impossible to increase in proportion to. That is, at the conventional rotational speed of double speed, the rotational speed of the rotor 15 is too high, and the time during which the pump chamber PR is located in the area (volume expansion area) of each of the suction ports 21 and 31 becomes too short. The pump chamber PR shifts to the region of the switching ports 23 and 33 (upstream of the volume reduction region) communicating with the discharge ports 22 and 32 without being able to suck in sufficient hydraulic oil. Therefore, even if the region of the switching ports 23 and 33 is shifted to, the pump chamber PR facing them is still in a negative pressure state, and hydraulic fluid can be discharged to the switching ports 23 and 33. It is impossible to increase the discharge amount as the engine speed R increases.

  Such a technical problem is particularly noticeable in a vane pump such as the oil pump 10. For example, in the case of a trochoid pump, the engine speed increases even after cavitation occurs. Since the discharge amount increases with this, the technical problem that the discharge amount is insufficient hardly occurs.

  Therefore, in the oil pump 10, when the discharge pressure P reaches the port switching oil pressure Pk corresponding to the discharge pressure at the engine speed Rk at which cavitation can occur due to the increase in the engine speed R, FIG. ), The urging force based on the internal pressure of the back pressure chamber 45 overcomes the set load Wk of the valve spring 44 in the control valve 40, and the valve element 43 is plugged against the set load Wk of the valve spring 44. By moving to the 42 side, the control valve 40 is switched from the first state to the second state. As a result, the first port 46 and the third port 48 communicate with each other through the annular space 50, and the switching ports 23 and 33 communicate with the switching ports 23 and 33 through the third communication hole 33 a and the second communication passage 39. Since the hydraulic oil is introduced from the suction ports 21 and 31, in addition to the suction ports 21 and 31, the switching ports 23 and 33 also function as suction ports.

  Thus, since each of the switching ports 23 and 33 functions as a suction port, the suction section (suction area) is expanded together with each of the suction ports 21 and 31, and the time during which the pump chamber PR is located in the suction area is long. Therefore, as described above, the hydraulic oil cannot be sufficiently supplied only in the area of each of the suction ports 21 and 31, and the negative pressure is also obtained on the upstream side of the volume reduction area (area of the switching ports 23 and 33). Sufficient hydraulic fluid can be supplied to the pump chamber PR that has been used, and the cavitation is suppressed.

  While the control valve 40 is switched from the first state to the second state, the communication amount between the first port 46 and the second port 47 in the control valve 40 is gradually reduced. As a result, the specific discharge amount (discharge amount per one rotation of the pump) decreases, so that even if the engine speed R increases, the increase amount of the discharge pressure P temporarily becomes almost flat, but the port switching is performed. After the completion, the discharge pressure P increases again in proportion to the engine speed R as the engine speed R increases (section d in FIG. 15).

  At this time, when the control valve 40 shifts to the second state, the discharge region is reduced, and each discharge port 22 is in a state where the volume of the pump chamber PR located in the discharge region is reduced. , 32, the hydraulic oil is discharged, so that the theoretical discharge amount is reduced as compared with the case where the control valve 40 is in the first state, but no cavitation occurs. The actual discharge amount will increase.

  Subsequently, when the engine speed R further increases and the discharge pressure P reaches the first operating oil pressure Ps set higher than the third engine demand oil pressure P3, as shown in FIG. Due to the urging force based on the internal pressure of 37, the cam ring 17 starts to move further in the concentric direction against the spring load Wx of the first spring 34. Then, the increase in the discharge amount is restricted as the eccentric amount of the cam ring 17 becomes smaller, and thereby the increase in the discharge pressure P due to the increase in the engine speed R is suppressed (section e in FIG. 15).

  Thus, in the oil pump 10, the cam ring 17 is controlled to swing by the springs 34, 35 and the control valve 40 so as to increase the discharge pressure P in a multi-step manner, so that the discharge pressure P is wasted. Without increasing, it is possible to obtain characteristics that are as close as possible to the required oil pressure (curve E) of the internal combustion engine (see FIG. 15).

  Moreover, in the case of the oil pump 10, the switching ports 23 and 33 are not simply connected to the suction ports 21 and 31 at a high speed when the discharge pressure P exceeds the port switching hydraulic pressure Pk. In order to suppress cavitation, a configuration in which more hydraulic oil is sucked in at the time of the high rotation, and an appropriate discharge amount is ensured by this, so that the pump 10 performs almost no wasteful work. There will be no deterioration of fuel consumption.

  As described above, according to the oil pump 10 according to the present embodiment, the high rotation speed such that the pump chamber PR located on the upstream side of the volume reduction region (the region of the switching ports 23 and 33) has a negative pressure. Occasionally, by switching the control valve 40, it is possible to supply hydraulic oil from each of the switching ports 23 and 33, thereby increasing the intake amount. Can be suppressed, and a sufficient discharge amount can be secured.

  In other words, in the case of the oil pump 10, the hydraulic oil can be sucked from the switching ports 23 and 33 at the time of the high rotation, so that the hydraulic oil can be sucked at the time of the high rotation ( (Suction area) is enlarged, and it is possible to secure a longer time (suction time) for the pump chamber PR to be located in the suction area, so that a sufficient suction amount can be secured and cavitation can be suppressed. Will be served.

  Further, in the oil pump 10, when the cam ring 17 moves in the concentric direction, a part of each of the suction ports 21 and 31 is covered by each side surface of the arm portion 17b. It becomes resistance, and the cavitation can be effectively suppressed.

  In addition, since the oil pump 10 is driven at a rotational speed twice that of the conventional (crankshaft rotational speed), it is possible to have the same discharge capacity as a conventional pump with half the conventional pump volume. As a result, the pump itself can be greatly reduced in size.

  In addition, in the present embodiment, since the existing balancer device 1 associated with the internal combustion engine is used to obtain the rotational driving force twice that of the prior art, a new driving source is not used and the cost is increased. It is also used for suppression of conversion.

  FIG. 17 shows a second embodiment of the oil pump according to the present invention, and the engine operating state is between the back pressure port 49 of the control valve 40 and the discharge ports 22 and 32 according to the first embodiment. Accordingly, a normally closed solenoid valve SV that operates based on the excitation current from the vehicle-mounted ECU 60 is provided, and the port switching control is electrically performed by the solenoid valve SV. FIG. 17A shows a state in which the solenoid valve SV is not energized with excitation current, and shows the first state according to the present invention. FIG. 17B shows a state in which the solenoid valve SV is energized with excitation current. And each represents the 2nd state concerning the present invention.

  That is, the solenoid valve SV is basically configured to be controlled with the rotational speed R of the internal combustion engine detected by a predetermined sensor or the like, and the engine rotational speed R is set to the predetermined rotational speed Rk (FIG. 15). In the state lower than the reference), the exciting current is not supplied from the ECU 60, the hydraulic oil guided to the back pressure port 49 is drained, and the discharge pressure P is not introduced into the back pressure chamber 45. As a result, the valve body 43 is pressed against the step 41c of the valve body 41 by the urging force of the valve spring 44, and the control valve 40 is maintained in the first state (FIG. 17 ( a)).

  On the other hand, when the rotational speed R of the internal combustion engine becomes equal to or higher than the predetermined rotational speed Rk, the solenoid valve SV is opened by energizing the exciting current from the ECU 60, and the back pressure chamber 45 has the port switching hydraulic pressure Pk or higher. The discharge pressure P is introduced. As a result, the valve body 43 moves to one end side (the plug 42 side) of the valve body 41 against the biasing force of the valve spring 44 with the introduction pressure, and the control valve 40 is moved to the second state. It will be switched (see FIG. 17B).

  As described above, in this embodiment, since the port switching control by the control valve 40 is electrically performed by using the solenoid valve SV, the discharge pressure (operating hydraulic pressure) is used as in the first embodiment. ) Is not affected by wear of each part of the pump 10 or an oil pressure change due to a change in the oil type, etc., so that the port switching control can be performed properly at all times. Can be more reliably suppressed.

  In the present embodiment, the pressure introduced into the back pressure chamber 45 of the control valve 40 by interposing the solenoid valve SV between the back pressure port 49 of the control valve 40 and the discharge ports 22 and 32. However, in addition to this configuration, the valve body 43 of the control valve 40 may be directly driven by the solenoid valve SV. Even in this case, the above-described effects can be obtained.

  FIG. 18 shows a third embodiment of the oil pump according to the present invention. The solenoid valve SV according to the second embodiment is changed to a linear solenoid valve RSV, and the linear solenoid valve RSV and each discharge port 22 are changed. 32, an orifice 61 is provided, and the pressure introduced into the back pressure chamber 45 is controlled by controlling the drain discharge amount with the linear solenoid valve RSV. 18A shows a state in which the drain discharge amount at the linear solenoid valve RSV is maximized and is the first state according to the present invention. FIG. 18B shows a drain discharge amount at the linear solenoid valve RSV is zero. And the second state according to the present invention is shown.

  That is, the linear solenoid valve RSV is accommodated in a substantially cylindrical shape with an opening formed at one end and closed at the other end, and slidable along the axial direction on the inner periphery of the valve body 51. The valve body 51 is formed by the valve body 52 having first and second land portions 52a and 52b formed in sliding contact with the inner peripheral surface of the valve body 51 at each end thereof, and the second land portion 52b of the valve body 52. And is attached to one end opening of the valve body 51 for energization. Along with this, the rod 54 b is advanced to mainly consist of an electromagnetic unit 54 that moves the valve body 52 in the axial direction to the other end side of the valve body 51 against the urging force of the valve spring 53.

  The valve body 51 is connected to the peripheral wall with an IN port 51a communicating with the switching ports 23 and 33, an OUT port 51b connected to a back pressure port 49, and the suction ports 21 and 31 or the outside. A drain port 51c for discharging hydraulic oil in the annular space 56, which will be described later, is formed therethrough.

  The valve body 52 is formed with a reduced diameter at its axial intermediate portion, and an annular space 56 is formed between the valve body 51 and the land portions 52a and 52b. The ports 51a to 51c communicate with each other. Then, when the drain port 51c is opened and closed by the second land portion 52b of the valve body 52, the hydraulic pressure that is led from the annular space 56 to the back pressure port 49 via the OUT port 51b is controlled.

  As is well known, the electromagnetic unit 54 is provided with a coil unit 54a in which a coil is wound around a bobbin and a yoke is externally fitted to the bobbin, and on the inner peripheral side of the coil unit 54a so as to be movable back and forth in the axial direction. An armature (not shown) made of a magnetic material and a rod 54b coupled to the armature and moving forward and backward with the armature according to the energization state of the coil.

  With this configuration, the linear solenoid valve RSV is in a state in which the maximum excitation current is supplied from the ECU 60 and the rod 54b is advanced to the maximum when the rotational speed R of the internal combustion engine is lower than the predetermined rotational speed Rk. The discharge amount of 51c becomes the maximum. As a result, the hydraulic pressure sufficient to operate the valve body 43 against the biasing force of the valve spring 44 is not introduced into the back pressure chamber 45 of the control valve 40, and the valve body is driven by the biasing force of the valve spring 44. 43 is pressed against the step portion 41c of the valve body 41, and the control valve 40 is maintained in the first state (see FIG. 18A).

  On the other hand, when the rotational speed R of the internal combustion engine increases and becomes equal to or higher than the predetermined rotational speed Rf corresponding to the first operating oil pressure Pf, the exciting current from the ECU 60 is gradually reduced as the engine rotational speed R increases. As the opening area of the port 51c gradually decreases, the internal pressure of the back pressure chamber 45 gradually increases. When the internal pressure of the back pressure chamber 45 reaches the port switching oil pressure Pk, the valve body 43 is connected to the valve body 41. The movement to the one end side (plug 42 side) starts and the port switching control is started. Eventually, when the excitation current from the ECU 60 is minimized and the drain port 51c is completely closed, the control valve 40 is completely shifted from the first state to the second state (FIG. 18). (See (b)).

  As described above, in the present embodiment, the linear solenoid valve RSV is employed for controlling the pressure introduced into the back pressure chamber 45 of the control valve 40, so that the same operational effects as in the second embodiment can be achieved. Of course, the linear solenoid valve RSV, unlike the solenoid valve SV capable of only ON-OFF control, can perform the port switching gradually, so that the discharge pressure based on the change in the discharge amount at the time of the port switching. It serves to suppress fluctuations.

  In this embodiment as well, the linear solenoid valve RSV is introduced between the back pressure port 49 of the control valve 40 and the discharge ports 22 and 32 to introduce the control valve 40 into the back pressure chamber 45. Although the pressure is controlled, the valve body 43 of the control valve 40 may be directly driven by the linear solenoid valve RSV in addition to the above structure. Even in this case, the above-described effects can be obtained. This is the same as in the second embodiment.

  The present invention is not limited to the configuration of each of the above-described embodiments. For example, the engine required oil pressures P1 to P3, the first and second operating oil pressures Pf and Ps, and the port switching oil pressure Pk are determined by the oil pump 10. It can be freely changed according to the specifications of an internal combustion engine, a valve timing control device, and the like of the mounted vehicle.

  Further, the drive source of the oil pump 10 is not necessarily required to use the balancer device 1, and other devices may be used as long as they can be driven at a higher rotational speed than the rotational speed of the internal combustion engine. Is possible. Further, the other device may be an existing device in the vehicle body or may be newly added.

  In each of the above embodiments, the suction region is expanded by connecting the switching ports 23 and 33 to the suction ports 21 and 31 through the control valve 40. Rather than uniformly expanding the suction area by combining the ports, the suction area at the time of high rotation may be secured by continuously expanding the suction area.

  That is, the discharge ports 22 and 32 and the switching ports 23 and 33 are configured as a series of discharge ports, and a pair of partition walls that partition the series of discharge ports and the suction ports 21 and 31. One partition wall (for example, the lower partition wall in FIG. 7) facing the pump chamber PR having the larger internal volume is separated from the members 11 and 12, which are separate from the members 11 and 12. The partition area may be configured as a relatively movable partition member, and the partition area may be moved in the circumferential direction to continuously change the distribution of the suction area and the discharge area, thereby expanding the suction area.

  In this way, by configuring the distribution of the suction area and the discharge area to be continuously variable, not only can the intake volume be secured at the time of high engine rotation, but also a rapid pump discharge volume based on the expansion / contraction of the suction / discharge area It is possible to suppress the change in the pressure and the smooth discharge amount control of the pump 10 is provided. Further, in the case of the above-described configuration, the suction / discharge area can be easily expanded and contracted only by moving the partition member, so that the response of the pump 10 can be improved.

  Further, in the second and third embodiments, the basic engine speed is described as an example of the control parameters of the solenoid valve SV and the linear solenoid valve RSV. However, in addition to the engine speed, for example, in the engine Control is performed according to the flowing coolant temperature or hydraulic oil temperature, or both, and the pressure in each switching port 23, 33 detected by providing a pressure sensor, a pressure switch, or the like in each switching port 23, 33. It is good as well.

  That is, since the occurrence of cavitation also affects the viscosity of the hydraulic oil, it is possible to more appropriately suppress the occurrence of the cavitation by setting the cooling water temperature or the hydraulic oil temperature, preferably both as control parameters. It becomes. Further, as described above, cavitation occurs when the internal pressure of each of the switching ports 23 and 33 becomes negative. Therefore, it is also possible to appropriately perform the port switching control based on the internal pressure of the switching ports 23 and 33. This is effective in suppressing cavitation.

  The technical ideas other than the invention described in the scope of claims ascertained from the respective embodiments will be described below.

(A) In the oil pump according to claim 1,
The pump structure is configured to vary the discharge amount according to the discharge pressure,
The oil pump, wherein the discharge pressure at which the valve body switches the first state to the second state is set lower than the discharge pressure at which the discharge amount per one rotation of the pump is the lowest.

  Cavitation occurs before the pump discharge pressure reaches the discharge pressure at which the pump's specific discharge amount reaches the minimum, so the switching pressure is set lower than the discharge pressure at which the specific discharge amount reaches the minimum. By doing so, it is possible to effectively suppress cavitation.

(B) In the oil pump according to (a),
The pump structure is driven by an internal combustion engine and has a rotor having a plurality of radial grooves on the outer peripheral side thereof, and is arranged on the outer peripheral side of the rotor so that the rotation center of the rotor and the center of the inner peripheral surface thereof are The cam ring is configured to be eccentric, and the vane is housed in the plurality of grooves so as to be able to protrude and retract, and each tip is in sliding contact with the inner peripheral surface of the cam ring,
An oil pump characterized in that the hydraulic oil chambers are separated by the port block, the rotor, the cam ring, and the vanes by disposing the port blocks on both axial sides of the cam ring.

(C) In the oil pump according to (b),
When the discharge pressure reaches the first pressure, the amount of eccentricity between the rotation center of the rotor and the center of the inner peripheral surface of the cam ring decreases, and the discharge pressure becomes a second pressure higher than the first pressure. The amount of eccentricity between the rotation center of the rotor and the center of the inner peripheral surface of the cam ring is further reduced,
The oil pump is characterized in that a discharge pressure at which the valve body switches the first state to the second state is set higher than the first pressure and lower than the second pressure.

  With this configuration, the function and effect of the present invention can be obtained while ensuring the function as a variable displacement pump. That is, since the discharge amount is reduced by switching from the first state to the second state, the swing timing of the cam ring is delayed, resulting in unnecessary work for the pump. . Therefore, if the above-mentioned switching is performed in a state where cavitation cannot occur, the cavitation reduction effect is not utilized, and conversely, the function as a variable displacement pump such as reducing the pump driving torque is wasted. Will be disturbed.

(D) In the oil pump according to (c),
The cam ring is configured such that the urging force of two types of springs having different spring loads acts on the cam ring,
When the discharge pressure reaches the first pressure, the cam ring resists the biasing force of a spring having a relatively small spring load among the two springs, and the center of rotation of the rotor and the center of the inner peripheral surface of the cam ring. Move in the direction that the amount of eccentricity decreases,
When the discharge pressure reaches the second pressure, the cam ring resists the biasing force of the spring having a relatively large spring load among the two springs, and the center of rotation of the rotor and the center of the inner peripheral surface of the cam ring. An oil pump configured to move in a direction in which the amount of eccentricity further decreases.

(E) In the oil pump according to (c),
The oil pump is configured such that when the amount of eccentricity between the rotation center of the rotor and the inner peripheral surface of the cam ring decreases, a part of the suction port is covered with a side surface of the cam ring.

  With this configuration, it is possible to reduce the amount of suction from the suction port, which is used to suppress the occurrence of cavitation.

(F) In the oil pump according to claim 1,
The oil pump is characterized in that the pump structure is driven at a higher speed than the rotational speed of the crankshaft of the internal combustion engine.

  As described above, when the pump having the same capacity is configured by increasing the rotational speed of the pump as compared with the conventional one, the pump can be made smaller than the conventional one.

(G) In the oil pump according to (f),
The oil pump is characterized in that the pump structure is driven at a rotational speed twice that of the crankshaft.

  In this way, by setting the number of rotations of the pump to be twice that of the conventional one, when a pump having the same capacity is configured, half the capacity of the conventional one is sufficient, and the pump can be greatly reduced in size.

(H) In the oil pump according to (g),
The pump structure is driven by a balancer device that reduces secondary vibration of an internal combustion engine.

  In this way, by using the existing device attached to the internal combustion engine, it is possible to obtain a rotational driving force twice that of the conventional one without using a new driving source, thereby suppressing an increase in cost. Provided.

(I) In the oil pump according to (h),
The oil pump is characterized in that the pump structure is incorporated in the balancer device.

  By integrating the pump and the balancer device in this manner, it becomes possible to mount the pump compactly on a vehicle or the like as compared with the case where the pump is separately arranged.

(J) In the oil pump according to claim 2,
The oil pump is characterized in that the control means is electrically controlled.

  By adopting such a configuration, it is not affected by oil pressure change due to oil pump wear or oil type change as in the case where the switching control is performed using the discharge pressure, so that the switching control is always performed appropriately. Is possible.

(K) In the oil pump according to (j),
The oil pump is characterized in that the control means is controlled according to the detected rotational speed of the internal combustion engine.

  Since the occurrence of cavitation depends on the rotational speed of the pump, the control according to the rotational speed of the internal combustion engine correlated with the rotational speed of the pump can appropriately suppress the occurrence of cavitation.

(L) In the oil pump according to (k),
The oil pump is characterized in that the control means is controlled in consideration of the detected temperature of the internal combustion engine.

  Since the occurrence of cavitation also affects the viscosity of the hydraulic oil, such a configuration makes it possible to more appropriately suppress the occurrence of cavitation.

(M) In the oil pump according to (j),
The oil pump is characterized in that the control means is controlled according to the detected pressure of the switching port.

  Cavitation occurs when the pressure in the switching port that opens to the upstream side of the area where the volume is reduced becomes negative. Therefore, by performing the switching control based on the pressure in the switching port, the occurrence of cavitation is appropriately suppressed. can do.

(N) In the oil pump according to claim 3,
An oil pump characterized in that when the hydraulic oil chamber upstream of the area where the volume is reduced has a negative rotation speed, the area for sucking the hydraulic oil is enlarged while the area for discharging the hydraulic oil is reduced. .

  By adopting such a configuration, it is possible to ensure a sufficient amount of inhalation in the inhalation process, and to suppress cavitation.

(O) In the oil pump according to (n),
The oil pump is characterized in that a region for sucking the hydraulic oil is continuously enlarged according to the rotation speed.

  By adopting such a configuration, it is possible to suppress a sudden change in the discharge amount based on the expansion / contraction of the suction / discharge region, and to provide smooth pump control.

(P) In the oil pump according to (o),
The port block is provided with a partition member for partitioning the suction port and the discharge port, and the partition member moves in accordance with the number of rotations, so that the hydraulic oil suction region is expanded. And oil pump.

  By adopting such a configuration, the suction / discharge area can be easily expanded / contracted as compared with a conventional pump in which the suction / discharge area is fixedly partitioned, and the response of the pump is improved.

10 ... Oil pump 11 ... Pump body (port block)
12 ... Cover member (port block)
15 ... Rotor (pump component)
16 ... Vane (pump component)
17 ... Cam ring (pump component)
18 ... Ring member (pump component)
21, 31 ... Suction ports 22, 32 ... Discharge ports 23, 33 ... Switching port 41 ... Valve body (valve housing hole)
43 ... Valve body 44 ... Valve spring (biasing member)
45 ... Back pressure port (Discharge pressure introduction port)
46a ... 1st port 47 ... 2nd port 48 ... 3rd port PR ... Pump chamber (hydraulic oil chamber)

Claims (3)

A pump structure for continuously increasing or decreasing the volume of the plurality of hydraulic oil chambers by being rotationally driven by the internal combustion engine;
A suction port that opens to an area in which the volume of the hydraulic oil chamber expands, a switching port that opens to an upstream side of an area in which the volume of the hydraulic oil chamber decreases, and an area in which the volume of the hydraulic oil chamber decreases. A port block having a discharge port opening on the downstream side of
A first port that communicates with the suction port, a second port that communicates with the switching port, and a third port that communicates with the discharge port are formed on the inner peripheral surface thereof, and at one end thereof, A valve housing hole provided with a discharge pressure introduction port for introducing a discharge pressure which is a hydraulic pressure in the discharge port;
A first state in which the second port and the third port communicate with each other by restricting communication between the first port and the second port by being slidably accommodated in the valve accommodating hole. And a valve body that switches between a second state that restricts communication between the second port and the third port by communicating the first port and the second port;
An urging member provided on the other end side of the valve accommodation hole, and urging the valve body toward one end side of the valve accommodation hole,
The pressure receiving area on one end side of the valve body on which the discharge pressure acts and the urging force of the urging member are from the first state when the pressure in the hydraulic oil chamber at which the switching port opens becomes negative. An oil pump, wherein the oil pump is set to be switched to the second state. A pump structure for continuously increasing or decreasing the volume of the plurality of hydraulic oil chambers by being rotationally driven by the internal combustion engine;
A suction port that opens to an area in which the volume of the hydraulic oil chamber expands, a switching port that opens to the upstream side of an area in which the volume of the hydraulic oil chamber decreases, and an area in which the volume of the hydraulic oil chamber decreases. A port block having a discharge port opening on the downstream side of
When the rotation speed is low such that the pressure in the hydraulic oil chamber where the switching port opens is positive, the switching port discharges hydraulic oil to the outside together with the discharge port, while the switching oil port opens the hydraulic oil chamber. An oil pump comprising: control means for controlling hydraulic oil to be supplied to the switching port from a low-pressure portion communicating with the suction port at a high rotation such that the pressure of the engine becomes negative. . A pump structure for continuously increasing or decreasing the volume of the plurality of hydraulic oil chambers by being rotationally driven by the internal combustion engine;
A port block having a suction port that opens to at least a region in which the volume of the hydraulic oil chamber increases, and a discharge port that opens to a region of the hydraulic oil chamber in which the volume decreases;
Control means for controlling the hydraulic oil to flow from the low pressure portion into the hydraulic oil chamber that becomes negative pressure when the pressure in the hydraulic oil chamber that is upstream of the area where the volume is reduced is a negative pressure; An oil pump comprising:

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