Variable displacement vane pump

A variable displacement vane pump that assuredly suppresses seizing of mutually contacting surfaces of a pressure plate and a rotor where the pressure plate is formed, at one side surface thereof that slidably contacts the rotor, with an annular lubricating groove. The lubricating groove is on a seal surface of the side surface at a position between arcuate back pressure grooves formed on the seal surface and a through opening which is formed in a center part of the pressure plate to receive the drive shaft. A radial width of the lubricating groove is set to a range from 10% to 25% of a radial width of the seal surface, and a distance from a center of the radial width of the lubricating groove to an inner cylindrical surface of the through opening is set to a range from 24% to 70% of the radial width of the seal surface.

TECHNICAL FIELD

The present invention relates to improvement in variable displacement vane pumps used for a power steering device of a motor vehicle.

BACKGROUND ART

For example, as a conventional variable displacement vane pump applied to a power steering device of a motor vehicle, a vane pump such as the vane pump disclosed in the latter mentioned patent reference 1 has been known.

The variable displacement vane pump of the reference comprises a cam ring that is swingably received in a receiving space formed in a front body, a rotor that is rotatably received in the cam ring and has vanes retractably and projectably received in radially extending slots formed in the rotor, a pressure plate that contacts an inside surface of the rotor, and a rear body that closes an open side of the receiving space of the front body.

The rotor has axially extending back pressure through bores respectively opened to the slots, and the pressure plate has, on an inside surface thereof facing the back pressure through bores, a generally arcuate back pressure groove that is connected a discharge chamber in which a pump discharge pressure is reserved. By introducing the pump discharge pressure to the back pressure bores through the back pressure groove, the vanes are forced to project from the corresponding slots and contact to an inner cylindrical surface of the cam ring, so that pump chambers are formed each being defined by mutually facing adjacent two vanes, an outer cylindrical surface of the rotor, the inner cylindrical surface of the cam ring, an outside surface of the pressure plate and an inside surface of the rear body.

The pressure plate and the rotor have, at respectively contacting surfaces thereof, radially spaced several annular zones each having a plurality of dimples depressed, each dimple having a generally arc-shaped cross section. The dimples function to lubricate the respective contact surfaces of the pressure plate and the rotor by temporarily reserving a high pressure oil that has entered thereinto from the back pressure groove of the pressure plate through a very fine clearance defined between the pressure plate and the rotor. With such dimples, undesired seizing of the mutually contacting surfaces of the pressure plate and the rotor is suppressed.Patent reference 1: Laid open Japanese Patent Application (Tokkai) 2000-337267

DISCLOSURE OF THE INVENTION

Task to be Solved by the Invention

In recent years, for much reducing an assist power of a steering action in the power steering device, variable displacement vane pumps having a high pump discharge pressure are in great demand.

However, when, in the conventional variable displacement vane pumps, the pump discharge pressure is set high, the pressure plate is pressed against the rotor with a higher pressing force, and thus, a seizing of the mutually contacting surfaces of the pressure plate and the rotor has not been suppressed by only providing the contacting surfaces with the above-mentioned dimples, which has been a severe problem.

The present invention is provided by taking the above-mentioned technical task into consideration and provides a variable displacement vane pump which assuredly suppresses the seizing between the mutually contacting surfaces of the pressure plate and the rotor.

Means for Solving the Task

According to one embodiment, there is provided a variable displacement vane pump which comprises a pump body that includes a front body having a receiving space defined therein and a rear body attached to the front body to close the receiving space; a drive shaft that passes through the pump body and rotatably supported in the same; a rotor that is mounted on an outer cylindrical surface of the drive shaft and received in the receiving space; a plurality vanes that are retractably and projectably received in a plurality of slots formed in the rotor in a manner to extend radially outward; a cam ring that is swingably arranged about the rotor to form a plurality of pump chambers, each being defined by adjacent two vanes, the rotor, and a part of the cam ring; a pressure plate that is arranged in a manner to be put between inside surfaces of the rotor and cam ring and a bottom surface of the receiving space and biased toward and pressed against the inside surface of the rotor to slidably contact with the same by a pump discharge pressure led from the bottom side of the receiving space; first and second fluid pressure chambers that are formed around the cam ring to control an eccentric amount of the cam ring; a pressure control means that controls the pressure in the first or second fluid pressure chamber; and an arrangement that includes an intake port that is provided by one of an inside surface of the rear body and the inside surface of the pressure plate facing the rotor and opened to a range where each pump chamber increases the volume; a discharge port that is provided by the above-mentioned selected inside surface and opened to a range where each pump chamber decreases the volume; an axially extending through opening that is formed in the pressure plate for receiving the drive shaft; a back pressure groove that is formed on the inside surface of the pressure plate at an area that slidably contacts the rotor to feed a pressurized fluid to bottom portions of the slots; a seal surface.

According to the invention, because the lubricating groove is constructed to the satisfy the above-mentioned condition, the mutually contacting surfaces of the rotor and pressure plate can be effectively lubricated even when the discharging pressure of the pump is set to a marked level. Thus, seizing between the rotor and the pressure plate is assuredly suppressed while suppressing lowering of the sealing performance of the seal surface.

According to the embodiment of paragraph [0009], characterization is so made that a depth of the lubricating groove is set to 25% or more of the radial width of the lubricating groove.

According to the invention, because the depth of the lubricating groove is set to satisfy the above-mentioned condition, it is possible to feed the lubricating groove with a large quantity fluid, and thus increase in lubrication performance of the lubricating groove is achieved. With this, the seizing of the rotor and pressure plate is much assuredly suppressed.

According to the embodiment of paragraph [0009], characterization is so made that the radial width of the lubricating groove is set to a range from 15% to 20% of the radial width of the seal surface.

According to the invention, because the radial width of the lubricating groove is set to satisfy the above-mentioned condition, only actually needed degree of lubrication is obtained without enlarging the radial width of the lubricating groove unnecessarily, and thus, lubrication performance and sealing performance are both obtained at the same time. With this, an optimum lubrication operation is carried out by the mutually contacting surfaces of the rotor and pressure plate, and thus seizing between the rotor and the pressure plate is assuredly prevented.

According to the embodiment of paragraph [0009], characterization is so made that the distance from the center of the radial width of the lubricating groove to the inner cylindrical surface of the through opening is set to a range from 30% to 45% of the radial width of the seal surface.

According to the invention, because the distance from the center of the radial width of the lubricating groove to the inner cylindrical surface of the through opening is set to the above-mentioned condition, the radial positioning of the lubricating groove is not excessively displaced and thus, a suitable sealing surface is obtained from the seal surface while obtaining an actually needed lubrication on the mutually contacting surfaces of the rotor and pressure plate. With this, leakage of the working fluid from the back pressure groove is much more effectively suppressed.

BEST MODE(S) FOR CARRYING OUT THE INVENTION

In the following, embodiments of the present invention, which are variable displacement vane pumps according to the present invention, will be described in detail with reference to the accompanying drawings. It is to be noted that the embodiments or the variable displacement vane pumps of the invention are those that are applied to a power steering device of a motor vehicle, like in the above-mentioned conventional pump.

That is, as is seen fromFIGS. 7 and 8, the variable displacement vane pump comprises a pump body1that is provided by joining a front body2and a rear body3, an annular adapter ring4that is tightly put in a receiving space2aformed in the pump body1, an annular cam ring6that is received in an oval space formed in the adapter ring4in a manner to be swingable about a swing fulcrum pin5, and a rotor8that is rotatably received in the cam ring6and fixed to a drive shaft7that passes through the pump body1.

The cam ring6has a width in an axial direction that is slightly smaller than the adapter ring4and the cam ring6is arranged in the receiving space2awhile keeping an eccentric position relative to the rotor8. Furthermore, the cam ring6is arranged to separate a first fluid pressure chamber10aand a second fluid pressure chamber10bthrough the swing fulcrum pin5and a seal member9that is placed at a position opposed to the fulcrum pin5.

The rotor8is shaped like a disc and has a width in an axial direction that is generally the same as that of the cam ring6. Furthermore, the rotor8, more specifically, opposed side surfaces of the rotor8in an axial direction are put or sandwiched, together with the cam ring6, between the rear body3and a circular pressure plate11of sintered material that is set on a bottom part of the receiving space2aof the front body2keeping a slight clearance “C” between the pressure plate and the rotor8as is seen fromFIG. 2.

The rotor8is arranged to rotate in a direction (counterclockwise direction) as is indicated by an arrow inFIG. 9when the drive shaft7is driven by an engine (not shown). The rotor has in an outer cylindrical portion thereof a plurality of slots8athat extend radially outward and are arranged at equally spaced positions. In the slots8a, there are received vanes12that are projectable radially outward toward the inner cylindrical surface of the cam ring6. A radially inside end of each slot8ais integrally formed with a back pressure chamber8bthat has a generally cylindrical shape.

In a space provided between the cam ring6and the rotor8, there is defined a pump chamber13by two adjacent vanes12, and by swinging the cam ring6about the swing fulcrum pin5, the volume of the pump chamber13is increased or decreased.

In the second fluid pressure chamber10b, there is installed a compression coil spring14, so that the cam ring6is constantly biased toward the first fluid pressure chamber10a, that is, in a direction to maximize the volume of the pump chamber13.

Furthermore, as is seen fromFIGS. 3 and 8, an inside surface3aof the rear body3at the side of the rotor3in an intake range “A” where the volume of each pump chamber13gradually increases in response to rotation of the rotor8is formed with a generally arcuate first intake port15. The first intake port15is formed at its middle portion with a first intake opening15athat is opened to an intake passage16formed in the rear body3, so that the working fluid led into the intake passage16from a reservoir tank (not shown) through an intake pipe17is led to each pump chamber13through the first intake opening15a.

Furthermore, as is seen fromFIG. 7, at a generally middle position of the inside surface3aof the rear body3, there is formed a recess3bthat bears one end portion of the drive shaft7, and at a bottom part of the recess3b, there is formed a return passage18that is connected with the intake passage16. The return passage18is so constructed as to permit the working fluid that has been led into the recess3bafter coming through the slight clearance “C” defined between the inside surface3aof the rear body3and the outside surface8cof the rotor8at the side of the rear body3to return to the intake passage16thereby leading the working fluid to the first intake port15through the first intake opening15a.

While, as is seen fromFIGS. 3 and 7, an inside surface11aof the pressure plate11at the side of the rotor8in a discharge range “B” where the volume of each pump chamber13gradually reduces in response to rotation of the rotor8is formed with a generally arcuate first discharge port19and a plurality of discharge openings20that are connected to the first discharge port19. Pressurized fluid discharged from the pump chambers13is led through the first discharge port19and the discharge openings20into a discharge side pressure chamber21that is formed in a bottom part of the receiving space2aof the front body2, and the pressurized fluid is then led to a hydraulic power cylinder of a power steering device (not shown) through a discharge passage (not shown) formed in the pump body1.

The inside surface11aof the pressure plate11is formed at a portion facing the first intake port15of the rear body3with a second intake port22that has a substantial same shape as the first intake port15. The second intake port22is formed at a middle portion thereof with a second intake opening22athat is opened to a relief passage23formed in the front body2, so that the working fluid returned from a relief valve40of an after-mentioned fluid control valve30through the relief passage23is led to each pump chamber13through the second intake opening22a.

Furthermore, the inside surface3aof the rear body3is formed at a portion facing the first discharge port19of the pressure plate11with a second discharge port24that has a substantially same shape as the first discharge port19. From both ends of the second discharge port24, there extend respective narrow grooves25aand25beach being sufficiently small as compared with the second discharge port24, and the narrow grooves25aand25bextend in a circumferential direction to positions near the ends of the first intake port15, so that production of noises that would be caused by sudden pressure change in each pump chamber13can be suppressed.

As will be understood from the above description, by providing the respective inside surfaces3aand11aof the rear body3and the pressure plate11with the first and second intake ports15and22and the first and second discharge ports19and24in which each pair is arranged symmetrically in an axial direction, the pressure balance between axially opposed portions of each pump chamber13is kept.

As is shown inFIG. 7, the pressure plate11is formed at a center portion thereof with a through opening26through which the drive shaft7passes, and the front body2is formed at a bottom part of the receiving space2awith a shaft hole2bfor bearing the other end of the drive shaft7, the shaft hole2bextending coaxially with the through opening26. These through opening26and shaft hole2bhave each an inside diameter that is slightly larger than an outside diameter of the drive shaft7, and thus between each of inner cylindrical surfaces of the through opening26and the shaft hole2band an outer cylindrical surface of the drive shaft7, there is defined a cylindrical oil passage27into which the working fluid that has flowed out from the clearance “C” between the inside surface11aof the pressure plate11and the inside surface8dof the rotor8placed at the side of the pressure plate11is led.

The shaft hole2bis provided at a generally middle position in an axial direction with a seal member28of which inside surface is formed with an annular groove28a, so that a clearance between the inner cylindrical surface of the shaft hole2band the outer cylindrical surface of the drive shaft7is sealed. Furthermore, in the front body2, there is formed a return passage29of which one end is opened to the annular groove28aof the seal member28and of which other end is connected to the relief passage23, so that the working fluid led into the cylindrical oil passage27is returned to the relief passage23through the annular groove28athereby to re-lead the working fluid to the second intake port22through the second intake opening22a.

Furthermore, as is seen inFIG. 7, the front body2is provided at its upper inner portion with a flow control valve30that controls a discharged amount of the fluid from the pump, the valve being arranged to extend perpendicular to the drive shaft7. As is seen fromFIG. 8, the flow control valve30comprises a spool element32that is slidably received in a valve hole31formed in the front body2, a valve spring34that biases the spool element32leftward in the drawing thereby to contact the element32to a plug33of the valve hole31, a high pressure chamber35that is defined between the plug33and a leading end of the spool element32to receive therein a fluid pressure appearing at an upstream side of a metering orifice (not shown), that is, a pressurized fluid that has been led into the discharge side pressure chamber21and a medium pressure chamber36that installs the valve spring34and receives therein the fluid pressure appearing at a downstream side of the metering orifice, so that when a pressure difference between the medium pressure chamber36and the high pressure chamber35exceeds a predetermined value, the spool element32is moved rightward in the drawing against the biasing force of the valve spring34.

When the spool element32is at a left position inFIG. 8, the first fluid pressure chamber10ais in communication with a low pressure chamber37defined around the spool element32through a communicating passage38that communicates the first fluid pressure chamber10ato the valve hole31. Into the low pressure chamber37, there is introduced a low pressure from the intake passage16through a low pressure passage (not shown) branched from the intake passage16.

When the spool element32is moved rightward in the drawing due to the work of a pressure difference between the chambers35and36, the low pressure chamber37is gradually blocked and the first fluid pressure chamber10abecomes communicated with the high pressure chamber35thereby to receive therein a highly pressurized fluid. That is, into the first fluid pressure chamber10a, there are selectively introduced a fluid pressure of the low pressure chamber37and a fluid pressure appearing at an upstream side of the metering orifice.

While, as is shown inFIG. 2, the second fluid pressure chamber10bis communicated with the first intake port15through a communicating passage39that is formed in the inside surface3aof the rear body3and extends radially outward from a position near the second fluid pressure chamber10bof the first intake port15, so that the second fluid pressure chamber10bis forced to constantly receive a fluid pressure (low pressure) of the intake side.

When the pressure in the medium pressure chamber36reaches a predetermined value, that is, when the working pressure in the above-mentioned power cylinder reaches the predetermined value, the relief valve40installed in the spool element32is opened to release the pressurized fluid to the relief passage23.

Furthermore, as is seen fromFIGS. 3 and 7, the inside surface11aof the pressure plate11is formed, at a portion facing the back pressure chamber8bin the intake range “A”, with an arcuate first intake side back pressure groove41that has a predetermined length in a circumferential direction. The first intake side back pressure groove41is formed at both ends with respective communicating bores41athat are connected to the above-mentioned discharge side pressure chamber21, so that part of the pressurized fluid in the discharge side pressure chamber21is led to the respective back pressure chambers8bthrough the communicating bores41a.

Furthermore, the inside surface is formed, at a portion facing the back pressure chamber8bin the intake range “A”, with a first discharge side back pressure groove42that has substantially the same shape as the above-mentioned first intake side back pressure groove41, the first discharge side back pressure groove42being placed on a common imaginary circle at a position diametrically opposed to the position where the first intake side back pressure groove41is placed (symmetrically upper and lower positions inFIG. 3). In the first discharge side back pressure groove42, there is arranged an axially extending orifice42athat is connected with the discharge side pressure chamber21, so that the pump discharging pressure is led into the groove42through the orifice42a.

As is seen fromFIG. 3, on the inside surface11aof the pressure plate11, there is defined around the through opening26a generally circular seal surface43due to provision of the first intake side back pressure groove41and the first discharge side back pressure groove42, and on the seal surface43, there is formed a circular lubricating groove44for lubricating a contacting with the rotor8.

As is seen fromFIGS. 1 and 3, the lubricating groove44has a generally rectangular cross section when viewed in a lateral direction and extends continuously in a circumferential direction without cuts, so that the seal surface43is partitioned into an outside seal surface43aand an inside seal surface43b. The lubricating groove44is positioned in a zone of the seal surface43where a distance “L” from an inner cylindrical surface26aof the through opening26to a center “P” of a radial width of the lubricating groove44indicates 24% to 70% of a total radial width “W1” of the seal surface43. In the illustrated embodiment, the distance “L” of the center “P” of the width of the lubricating groove44is set to 30% to 45% of the total width “W1”, which was proved by satisfied results of an after-mentioned experiment.

A radial width “W2” of the lubricating groove44is set to 10% to 25% of the total radial width “W1” of the seal surface43, and the depth “D” is set to a range that is larger than 25% of the total width “W1” of the seal surface43. In the illustrated embodiment, the radial width “W2” of the lubricating groove44is set to 15% to 20% of the total radial width “W1” of the seal surface43, which was proved by satisfied results of an after-mentioned experiment.

As is seen fromFIGS. 4 and 7, a bottom surface of the receiving space2aof the front body2is formed with an annular seal holding groove45that has a generally mushroom-shaped cross section when viewed in an axial direction. As is indicated by a phantom line inFIG. 4, with respect to the outside surface11bof the bottom side of the receiving space2aof the pressure plate11, the seal holding groove45has an inside part that extends around a lower half of the through opening26and an outside part of which opposed ends radially outward extend from opposed peripheral ends of an upper half of the through opening26and extend toward each other to enclose a center portion of the second intake port22. The seal holding groove45has a seal member46of rubber material tightly fixed thereto.

As is seen fromFIG. 4, at the outside surface11bof the pressure plate11, due to presence of the seal member46, inside and outside portions of a peripheral area of the second intake opening22aform respectively a low pressure zone “Lp” that communicates with the intake side and a high pressure zone “Hp” that communicates with the discharge side, and as is seen fromFIG. 7, to the low pressure zone “Lp” enclosed by the seal member46, there is applied the fluid pressure (low pressure) that has been led from the relief passage23, and to the high pressure zone “Hp” provided around the seal member46, there is applied the fluid pressure (high pressure) that has been led from the discharge side pressure chamber21.

While, as is seen fromFIG. 2, the inside surface3aof the rear body3is formed, at a position that faces the first intake side back pressure groove41of the pressure plate11, with a second intake side back pressure groove47that has substantially the same shape as the above-mentioned first intake side back pressure groove41. Furthermore, the rear body3is formed, at a position that faces the first discharge side back pressure groove42, with a second discharge side back pressure groove48that has substantially the same shape as the above-mentioned first discharge side back pressure groove42, the second discharge side back pressure groove48being placed generally symmetrical to the second intake side back pressure groove47(symmetrically upper and lower positions inFIG. 2). The second intake side back pressure groove47and the second discharge side back pressure groove48have respective ends that are connected through communicating grooves49aand49bwhose depth is small as compared with the grooves47and48.

On the inside surface3aof the rear body3, there is formed a generally circular seal surface50around the recess3b, which is defined by the second intake side back pressure groove47, the second discharge side back pressure groove48and the communicating groves49aand49b. The seal surface50is formed, at a portion facing the lubricating groove44, with a lubricating groove51that has substantially the same shape as the lubricating groove44.

In the following, unique operation of the variable displacement vane pump according to the embodiment will be described with reference toFIG. 2.

Upon operation, in the variable displacement vane pump, the pressure plate11is biased toward the rotor8by the pump discharging pressure, so that the inside surface11aof the pressure plate11is entirely in contact with the inside surface8dof the rotor8. Under this condition, the above-mentioned clearance “C” is produced between the inside surface11aof the pressure plate11and the inside surface8dof the rotor8, so that the pressure plate11is forced to have its center area maximally projected thereby causing a peripheral area of the through opening26to be highly pressed by the inside surface8dof the rotor8.

However, since the seal surface43that would be easily subjected to unbalanced wear when the inside surface11aof the pressure plate11slidably contacts the rotor8is formed with the lubricating groove44that fulfills the above-mentioned dimensional condition, the following desired fluid flow is made. That is, as is indicated by a broken line inFIG. 1, the pressurized fluid in the first intake side back pressure groove41and the first discharge side back pressure groove42is forced to flow to the outside seal surface43athrough the clearance “C”. Then, the pressurized fluid led between the outside seal surface43aand the rotor8is led into the lubricating groove44while lubricating contracting portions between outside seal surface43aand the inside surface8dof the rotor8.

Then, the pressurized fluid led into the lubricating groove44is temporarily kept therein and then led toward the inside seal surface43bfrom the lubricating groove44. Then, the pressurized fluid led between the inside seal surface43band the rotor8is led to a radially inner side of the pressure plate11, that is, to the above-mentioned cylindrical oil passage27while lubricating contacting portions between the inside seal surface43band the inside surface8d. Like this, the pressurized fluid led into the cylindrical oil passage27is led into the relief passage23through the annular groove28aof the seal member28and the return passage29, and returned back to the intake side pump chamber13through the second intake opening22a.

As is described hereinabove, by providing the seal surface43of the inside surface11aof the pressure plate11with the above-mentioned lubricating groove44, the above-mentioned lubrication operation is achieved. In particular, by setting the distance “L” to the center “P” of the radial width of the lubricating groove44, the radial width “W2” and the depth “D” to the above-mentioned ranges, seizing of an outer peripheral area of the through opening26at a side of the inside surface11aof the pressure plate11is assuredly suppressed due to an excellent lubrication operation. This excellent lubrication operation was proved by the results of an endurance test of a pump device, which will be described in the following.

FIG. 5is a graph showing the results of the endurance test in which the distance “L” to the center “P” of the radial width of the lubricating groove44and the radial width “W2” were randomly varied. In the graph, cases wherein the outer peripheral area of the through opening26at the side of the inside surface11awas not subjected to seizing are judged or indicated by “O”, while cases wherein the outer peripheral area was subjected to seizing are judged or indicated by “X”.

That is, in case wherein the distance “L” to the center “P” of the radial width of the lubricating groove44was set to 24% or less or 70% or more of the radial width “W1” of the seal surface43, the lubricating groove44was displaced too much in a radial direction, and thus, the seal area of the outside seal surface43aor the inside seal surface43bbecame so small resulting in that the sealing performance of one of the seal surfaces43aand43bwas excessively lowered and thus the lubrication operation of the lubricating groove44was insufficient.

While, in case wherein the distance “L” to the center “P” of the radial width of the lubricating groove44was set to a range from 24% to 70% of the radial width “W1” of the seal surface43, satisfied lubrication operation of the lubricating groove44was achieved. Particularly in case wherein the distance “L” was set to a range from 30% to 45% of the radial width “W1”, excessive displacement of the radial positioning of the lubricating groove44was not induced and thus suitable seal areas of the outside seal surface43aand the inside seal surface40bwere obtained resulting in excellent seal performance and satisfied lubrication operation.

While, in case wherein the radial width “W2” of the lubricating groove44was set to 10% a or less of the radial width “W1” of the seal surface43, the radial width W2” became too small and thus the groove could not hold the fluid sufficiently resulting in that a satisfied lubrication operation of the lubricating groove44was not obtained. While, in case wherein the radial width “W2” of the lubricating groove44was set to 25% or more of the radial width “W1” of the seal surface43, the seal area of the seal surface43became excessively small and thus the sealing performance of the seal surface43was excessively lowered resulting in unsatisfied lubrication operation of the lubricating groove44.

While, in case wherein the radial width “W2” of the lubricating groove44was set to a range from 10% to 25% of the radial width “W1” of the seal surface43, satisfied lubrication operation of the lubricating groove44was obtained. As is shown by a slashed portion ofFIG. 5, in case wherein the radial width “W2” was set to a range from 15% to 20% of the radial width “W1”, a suitable lubrication amount was obtained without increasing the radial width of the lubricating groove44to an unnecessarily large degree and thus the lubrication performance and the sealing performance were obtained at the same time inducing excellent lubrication operation.

By the above-mentioned experiment, it became apparent that, for obtaining a satisfied lubrication operation by the lubricating groove44, the distance “L” to the center “P” of the radial width of the lubricating groove4and the radial width “W2” are arranged within a zone “G” enclosed by a thicker line ofFIG. 5.

FIG. 6is a graph depicting the results of an endurance test wherein in the zone “G”, only the depth “D” of the lubricating groove44was changed while fixing the above-mentioned parameters “L” and “W2” to certain values. Like in the above-mentioned experiment, cases in which no seizing appeared on the sliding surface of the pressure plate11are indicated or judged by “O”, while cases in which seizing appeared on the sliding surface are indicated or judged by “X”.

That is, in case wherein the depth “D” of the lubricating groove44was set to 25% of the radial width “W2” of the lubricating grove44, the depth of the lubricating groove44became too small and thus the groove failed to hold therein a satisfied amount of fluid resulting in failure in exhibiting a satisfied lubrication operation of the lubricating groove44.

While, in case wherein the depth “D” of the lubricating groove44was set to 25% or more of the radial width “W2” of the lubricating groove44, satisfied lubrication operation of the lubricating grove44was obtained. As a result, it became apparent that the range of the depth “D” that assures a satisfied lubrication operation by the lubricating groove44was arranged above the thicker line ofFIG. 6, that is, a range larger than 25% of the radial width “W2”.

Accordingly, in the above-mentioned embodiment, by providing the seal surface43of the pressure plate11with the lubricating groove44that satisfies the above-mentioned set conditions of the radial position “L” and radial width “W2”, undesired seizing of the pressure plate11to the rotor8is assuredly prevented even if the pump discharging pressure is set high, because the outer peripheral part of the through opening26of the inside surface11aof the pressure plate11, which is strongly pressed against the rotor8, is effectively lubricated while effectively lowering deterioration of the sealing between the inside surface8dof the rotor8and the seal surface43.

Particularly, when the distance “L” to the center “P” of the radial width of the lubricating groove44is set to the range from 30% to 45% of the radial width “W1” of the seal surface43and at the same time the radial width “W2” of the lubricating groove44is set to the range from 15% to 20% of the radial width “W1” of the seal surface43, the lubricating groove44is prevented from taking an excessively displaced radial positioning and the seal surfaces43aand43bcan obtain suitable sealing surfaces, and thus, satisfied lubrication is obtained by only an amount of fluid actually needed without increasing the radial width of the lubricating groove44to an unnecessarily large degree. Accordingly, suppressing of lowering in pumping efficiency which would be inevitably caused by leakage of the pressurized fluid and lubrication between respective sliding surfaces of the rotor8and pressure plate11are most effectively achieved.

Furthermore, by setting the depth “D” of the lubricating groove44to a range that is larger than 25% of the radial width “W1” of the seal surface43, the lubricating groove44can hold a larger amount of pressurized fluid and thus the lubricating performance of the lubricating groove44can be increased, which assuredly prevents undesired seizing of the pressure plate11to the rotor8.

Furthermore, at the inside surface3aof the rear body3, substantially same operation as the above-mentioned lubrication operation at the inside surface11aof the pressure plate11is obtained since the clearance between the seal surface50and the outside surface8cof the rotor8is lubricated by the pressurized fluid that has leaked thereto through the lubricating groove51from the second intake side back pressure groove47and second discharge side back pressure groove48, and the pressurized fluid is returned back to the intake side from the recess3bthrough the return passage18.

Furthermore, by separating the outer peripheral part of the second intake opening22aof the outside surface11bof the pressure plate11into inside and outside areas with usage of the seal member46, the inside (viz., the side facing the rotor8) of an upper half of the pressure plate11is applied with the intake pressure and at the same time the outside (viz., the side facing the bottom portion of the receiving space2a) of the same is applied with a low-pressurized working fluid that has returned from the relief valve40and the return passage29, and thus the fluid pressure applied to the axially opposed both sides11aand11bcan be balanced. That is, undesired phenomenon wherein the upper half portion of the pressure plate11is deformed or depressed toward the rotor8is suppressed, and thus, increase in unbalanced wearing of the seal surface43, which would be caused by the high pressing of the upper half portion of the inside surface11aof the pressure plate11against the rotor8, is suppressed or lowered.

Furthermore, since the lubricating groove44has a generally annular cross section when viewed in an axial direction, the working fluid is permitted to make a recirculation flow in the lubricating groove44, and thus, the lubrication performance of the lubricating groove44is further increased.

Furthermore, the pressure plate11is produced by a sintered material and by using a die forming technique, the lubricating groove44can take various shapes freely.

Furthermore, since the sintered material is porous, the working fluid can be accumulated in the minute pores formed in the pressure plate11and thus the lubricating performance exhibited when the pressure plate11slidably contacts the rotor8is much increased.

FIG. 9shows a second embodiment of the present invention. The basic construction of this embodiment is the same as that of the above-mentioned first embodiment, and in the second embodiment, modification is made to the lubricating groove44of the first embodiment.

That is, at a part corresponding to the rest part of the low pressure zone “Lp” defined by the seal member46of the above-mentioned first embodiment, the lubricating groove44in the second embodiment is formed with a circumferentially extending narrower groove part52that has a radial width “W3” smaller than the radial width “W2” of the lubricating groove44.

Accordingly, in the second embodiment, a portion where both the side surfaces11aand11bof the pressure plate11are pressed by the discharge side fluid pressure (higher pressure), that is, a portion of which deformation tends to increase upon appearance of a pressure difference between the side surfaces11aand11bachieves a satisfied lubrication performance at a portion that contacts the rotor8, and a portion that is pressed by the intake side fluid pressure (lower pressure), that is, a portion of which deformation does not increase so large even upon appearance of the pressure difference between the side surfaces11aand11bexhibits a satisfied sealing performance to the mutually contacting surfaces of the pressure plate11and the rotor8in the low pressure zone “Lp” due to provision of the narrower groove part52.

Accordingly, also in the second embodiment, substantially same operation effects as those of the first embodiment are obtained. In addition to this, at a portion that needs a heavy lubrication, seizing is suppressed due to a satisfied lubrication at the mutually contacting surfaces and at a portion that needs only a light lubrication, leakage of the pressurized fluid from the back pressure grooves41and42is suppressed due to the increased sealing between the mutually contacting surfaces, so that the lubrication of the mutually contacting surfaces and suppressing of lowering of pumping effect which would be caused by the fluid leakage are both achieved effectively.

FIG. 10shows a third embodiment of the present invention. This third embodiment has no part that corresponds to the above-mentioned narrower groove part52of the first embodiment and has as a substitute for the part52an enlarged seal surface53, and in the third embodiment, the lubricating groove44has a generally C-shaped cross section when viewed in an axial direction.

That is, in the pressure plate11, only portions where lubrication is absolutely necessary are provided with the lubricating groove44, and portions where the lubrication is not absolutely necessary are formed with the enlarged seal surface53, so that the mutually contacting surfaces of the pressure plate11and rotor8are satisfied with both the lubrication and sealing performance. Thus, the lubrication of the mutually contacting surfaces and suppressing of lowering of pumping effect which would be caused by the leakage of the pressurized fluid are both achieved effectively.

FIG. 11shows a fourth embodiment of the present invention. The basic construction of this embodiment is the same as that of the above-mentioned first embodiment, and in this fourth embodiment, modification is made to the lubricating groove44of the first embodiment.

That is, at a part corresponding to the rest part of the low pressure zone “Lp” defined by the seal member46of the above-mentioned first embodiment, the lubricating groove44of the fourth embodiment is formed with a straight groove part54that extends leftward and rightward inFIG. 11and has a radial distance “L1” smaller than the distance “L” to the center “P” of the radial width of the lubricating groove44.

Accordingly, also in this embodiment, substantially same operation effects as those of the first embodiment are obtained. In addition to this, due to provision of the straight groove part54in the low pressure zone “Lp” of the lubricating groove44, the outside seal surface43ais expanded in the low pressure zone “Lp” where axial deformation of the pressure plate11is not easily carried out, and thus leakage of the pressurized fluid from the first intake side back pressure groove41is reduced thereby providing the lubricating groove44with a suitable lubrication performance as well as a suitable sealing performance.

In the following, technical thoughts possessed by the above-mentioned various embodiments, will be described.

A variable displacement vane pump which is characterized in that the lubricating groove has a generally arcuate cross section when viewed in a lateral direction.

According to the invention, since the lubricating groove is so constructed as to have a generally arcuate cross section when viewed in a lateral direction, it is possible to reduce a flow resistance produced when a fluid flows through the lubricating groove, and thus, improvement in lubrication of the working fluid in the lubricating groove is obtained.

A variable displacement vane pump according to paragraph [0077], which is further characterized in that the lubricating groove is constructed to have a generally annular cross section when viewed in an axial direction.

According to the invention, since the lubricating groove is constructed to have a generally annular cross section when viewed in an axial direction, it is possible to circulate the working fluid in the lubricating groove, and thus, the lubrication performance of the lubricating groove is much more increased.

A variable displacement vane pump according to paragraph [0077], which is further characterized in that the lubricating groove is formed on the pressure plate at a portion of the pressure plate which is largely deformed when a pump discharge pressure is applied to the pressure plate, and at other portion of the pressure plate which is not so largely deformed, there is formed a seal surface that carries out a sealing against the inside surface of the rotor.

According to the invention, the lubricating groove is formed on the pressure plate at a portion that is largely deformed upon application of the pump discharge pressure to the pressure plate, and at a portion that is not largely deformed, there is formed a seal surface that carries out a sealing against the inside surface of the rotor. With this, by the lubricating groove, lubrication performance and sealing performance are both assured at the same time.

A variable displacement vane pump according to paragraph [0081], which is further characterized in that the seal surface is formed on the side of the discharge port.

According to the invention, since, at the side of the discharge port, both of the axially opposed portions of the pressure plate show a higher pressure and thus respective pressures at such opposed portions are kept balanced, the axial deformation of the pressure plate upon application of pressurized fluid thereto is relatively small. Accordingly, by providing the discharge port side with the above-mentioned seal surface, leakage of the working fluid from the back pressure groove can be suppressed.

A variable displacement vane pump according to paragraph [0081], which is further characterized in that the lubricating groove is positioned at a portion of the pressure plate where a pressure difference between axially opposed portions of the pressure plate is large, and at a portion of the pressure plate where the pressure difference is small, there is provided a seal surface that effects a sealing against the inside surface of the rotor.

According to the invention, at a portion where the pressure different between the axially opposed portions of the pressure plate is large, the deformation of the portion is increased since the pressure plate is biased in one axial direction by the pressure difference. Thus, by providing the portion with the lubricating groove, seizing between the pressure plate and rotor is effectively suppressed. While, at the other portion where the pressure difference is small, by providing the other portion with the seal surface, the leakage of the working fluid from the back pressure groove is suppressed while preventing the seizing between the pressure plate and the rotor.

A variable displacement vane pump according to paragraph [0081], which is further characterized in that the pressure plate is provided, at a side thereof facing the receiving space of the front body, with a seal member by which a high pressure portion and a low pressure portion are partitioned, and in that the seal surface is arranged at the low pressure portion defined by the seal member.

According to the invention, since, at the low pressure portion defined by the seal member, the axially opposed sides of the pressure plate show a low pressure keeping a balanced pressure condition therebetween, the axial deformation of the pressure plate is small. Accordingly, by providing the low pressure portion with the above-mentioned seal surface, the leakage of the working fluid from the back pressure groove is suppressed.

A variable displacement vane pump according to paragraph [0077], which is further characterized in that the pressure plate is produced by using a die forming technique.

According to the invention, since the pressure plate is produced by using the die forming technique, the shape of the lubricating groove can be freely set.

A variable displacement vane pump according to paragraph [0089], which is further characterized in that the pressure plate is produced by a sintered material.

According to the invention, since the sintered material is porous, the working fluid can be accumulated in the minute pores formed in the pressure plate, lubrication performance of the pressure plate relative to the rotor is further increased.

A variable displacement vane pump according to paragraph [0089], which is further characterized in that the pressure plate is produced by a die-cast aluminum material.

According to the invention, since the pressure plate is produced by the die-cast aluminum material, reduction in weight of entire construction of the device is achieved. Furthermore, by adding a suitable amount of anti-wear additive to the material of the aluminum die-cast, the wear resistance of the pressure plate to the rotor is controlled.

A variable displacement vane pump according to paragraph [0089], which is further characterized in that the lubricating groove has differently shaped portions in a circumferential direction.

According to the invention, the amount of the lubrication fluid can be varied or adjusted in a circumferential direction in accordance with the need of the fluid by portions, such as a portion that needs a larger amount of the fluid and a portion that needs only a small amount of the fluid.

A variable displacement vane pump according to paragraph [0095], which is further characterized in that the lubricating groove is arranged at only one part of a circumferentially extending imaginary line.

According to the invention, at a portion that needs lubrication, there is provided the lubricating groove, and at a portion that does not need lubrication, there is provided the seal surface without providing the lubricating groove. With this, lubrication performance and sealing performance are achieved at the same time.

A variable displacement vane pump according to paragraph [0095], which is further characterized in that the lubricating groove has a radial width that varies in accordance with positions in a circumferential direction.

According to the invention, for the portion that needs a larger amount of lubrication, the lubrication performance can be increased by increasing the radial width of the lubricating groove, and thus seizing between the pressure plate and the rotor can be suppressed, and for the portion that needs only a small amount of lubrication, the sealing performance can be increased by reducing the radial width of the lubricating groove, and thus leakage of the working fluid from the back pressure groove can be suppressed.

A variable displacement vane pump according to paragraph [0095], which is further characterized in that the lubricating groove is so shaped that the distance to the center varies depending on a position taken in a circumferential direction.

According to the invention, since the lubricating groove is so shaped that the distance to the center varies depending on a position taken in a circumferential direction, the lubrication performance and the sealing performance are suitably achieved.

The present invention is not limited to the construction of the above-mentioned embodiments, and is applicable to a construction wherein the shape, size, etc., of the intake ports15and22, the discharge ports19and24and the back pressure grooves41,42,47and48are varied in accordance with the specification and size of the pump device.

Furthermore, by placing the narrow groove part52and the enlarged seal surface53of the above-mentioned second and third embodiments to their corresponding ranges “B”, that is, by symmetrically placing them at upper and lower positions inFIGS. 9 and 10, the seal member46and the seal holding groove45may be removed.

In such cases, the narrow groove part52and the enlarged seal surface53are arranged in a circumferential range of the discharge range “B” where the pump discharge pressure is applied to both the axially opposed side surfaces11aand11bof the pressure plate11to achieve a balanced pressure condition in the axial direction, and the lubricating groove44is arranged in only a circumferential range other than the discharge range “B”, where the pressure difference between the opposed side surfaces11aand11bis remarked. Thus, seizing of the sliding contact surface of the pressure plate11can be prevented suppressing the leakage of the pressurized fluid from the back pressure grooves41and42. Furthermore, since there is no need of providing the seal member46and the seal holding groove45, the cost of manufacturing can be reduced.

The lubricating groove44can have a generally arcuate cross section when viewed in a lateral direction. In this case, the flow resistance that appears when the pressurized fluid flows in the lubricating groove44can be reduced, and thus, the lubrication performance of the pressurized fluid flowing in the lubricating groove44can be increased.

Furthermore, the pressure plate11can be produced by a die-cast aluminum material. In this case, lightening of the entire construction of the pump device is achieved. Furthermore, by adding a suitable amount of anti-wear additive to the material of the aluminum die-cast, the wear resistance of the pressure plate to the rotor can be controlled.

DESCRIPTION OF THE REFERENCE NUMERALS