DESCRIPTION

Technical Field

The present invention generally refers to a volumetric rotary pump and, more particularly, to a volumetric vane pump with variable displacement. The invention finds particular application as an oil pump in motor vehicles.

Background Art

In the field of motor vehicles, volumetric rotary vane pumps with variable displacement are known, which are used, for example, as oil pumps. Such pumps generally comprise: a rotor, rotatably mounted around a longitudinal axis thereof and having a plurality of radial slots open towards the outside; a stator extending around the rotor and mounted eccentrically relative to the rotor; a plurality of vanes, each of which is slidably mounted in a respective slot of the rotor and has an end, distal to the rotor, arranged to come into contact with the inner wall of the stator; adjustment means for selectively controlling the eccentricity between rotor and stator and, therefore, the flow rate of the pump.

Volumetric rotary vane pumps with variable displacement of the type described above have wear problems because of the friction generated by the distal ends of the vanes sliding against the inner wall of the stator. Said friction also causes lower performance of the pump and thus higher energy consumption.

The aforesaid wear also depends on the various temperature conditions at which the pump operates: for this reason, the coefficients of thermal expansion of the materials of which the vanes and the stator (or the part thereof which is in contact with the vanes) are made should be as similar as possible.

In order to limit such problems, documentDE102011014591Al describes apump providing for the use of a steel ring arranged within the stator, between the inner wall of the stator and the distal ends of the vanes. The stator is made of a plastic material and the steel ring is co-molded within the stator.

The pump according to the aforesaid document, though limiting the problem of the wear of the stator by using the steel ring, does not suitably solve the problem of the friction between vanes and ring. The pump further has problems of loss of performance due to the different axial expansions of the stator and the ring co-molded therein, due to the different coefficients of thermal expansion of said components. The object of the present invention is to overcome the problems and limitations of prior art by providing a volumetric rotary pump that limits the wear between vanes and pump as well as the loss of performance.

This and other objects are achieved with the volumetric rotary pump as claimed in the appended claims.

Summary of Invention

The volumetric rotary vane pump according to the invention, preferably a variable displacement pump, comprises a stator which is housed in a cavity formed in a body of the pump and has an inner wall and an outer wall. A preferably cylindrical pumping chamber of the pump is defined by the inner wall of the stator, a bottom of the body of the pump and a pump lid.

Inside the pumping chamber, a suction inlet and a discharge outlet are preferably provided on the bottom of the pump body delimiting the pumping chamber.

The pump further comprises a vane rotor which is rotatably mounted within the pumping chamber and arranged eccentrically thereto, said vane rotor being configured to rotate around a rotation axis thereof by being driven by a driving shaft.

The pump further comprises a ring rotatably mounted within the pumping chamber and having an outer surface, facing the inner wall of the stator, and an inner surface facing the rotor. A thin annular interstice is provided between the inner wall of the stator and the outer surface of the ring.

A pressure chamber is also located within the cavity of the body of the pump, between an inner wall of the body of the pump and an outer wall of the stator. Said pressure chamber is suitable for containing a pressurized fluid produced by the pump. The pressure that is created in said pressure chamber is adapted to exert a thrust onto the stator, to make it rotate around a pivot axis thereof or translate, in order to vary the eccentricity between the pumping chamber and the rotor and, thus, the displacement of the pump.

The rotor comprises a plurality of slots, typically radial, open towards the outside of the rotor and having a blind bottom proximal to the rotation axis of the rotor. Each of the slots is adapted to slidably accommodate a respective vane of the rotor, each vane having a respective first end adapted, in use, to contact, in a tightly sealed manner, the inner surface of the ring, and a second end received in the blind bottom of the respective slot.

The ring is configured to be brought into rotation, in use, in the same direction as the direction of rotation of the rotor by the interaction between the first ends of the vanes and the inner surface of the ring. Accordingly, advantageously, the relative sliding between the vanes and the ring is reduced, if not eliminated, thereby reducing the friction-related wear between vanes and ring.

Preferably, the pump further comprises a first duct, formed in the stator and extending through the stator from the outer wall thereof to the inner wall thereof, thus connecting the pressure chamber to the interstice provided between the inner wall of the stator and the outer surface of the ring. Said first duct is adapted to allow the passage of pressurized fluid from the pressure chamber to said interstice, thus creating a bearing means between the inner wall of the stator and the outer surface of the ring. This meatus makes it possible to reduce the friction between the inner wall of the stator and the outer surface of the ring.

Preferably, said first duct opens into the interstice by exiting from the inner wall of the stator along a direction such that a flow of pressurized fluid reaching the interstice through the first duct has the same direction of rotation as the direction of rotation of the rotor, so as to impart a tangential thrust to the ring in the same direction in which the ring is brought into rotation by the interaction with the first ends of the vanes, thus promoting such rotation.

Preferably, the first duct exits from a point on the inner wall of the stator along a direction that is substantially tangent to said inner wall at said point. Alternatively, the first duct exits from a point on the inner wall of the stator along a direction that forms, with the direction tangent to the inner wall of the stator at said point of exit of the first duct, an angle greater than 0° and smaller than 90°, so as to ensure a tangential thrust, which again is in the same direction in which the ring is brought into rotation, though smaller, at the same fluid pressure, than the thrust generated when the first duct exits tangentially from the inner wall of the rotor.

Preferably, the ring and the vanes are made of metal, such as steel. In this way, the vanes and the ring exhibit greater resistance to wear. The stator is preferably made of a thermoset resin, possibly enriched with mineral fibers or carbon fibers in order to further reduce friction and the coefficient of thermal expansion of the resin, which is known to be higher than that of metals, or the stator is made of graphite or carbon- graphite, possibly impregnated with other materials. Thanks to the fact that the vanes and the ring are both made of metals and therefore have the same or similar friction coefficient, the contact-related wear problems are reduced. In addition, the stator, not being constrained to the rotating ring, can be dimensioned, axially and radially, in such a way as to achieve minimum axial clearance and maximum radial clearance, by virtue of the coefficient of thermal expansion, at high temperatures (at which the pump usually operates), thus preserving free movement between the stator, the ring and the cavity of the body of the pump, at the lowest temperatures of use, thus optimizing the volumetric efficiency of the pump in the thermal field of operation.

Preferably, the pump stator further comprises a second duct and a groove formed on the inner wall of the stator.

The second duct extends through the stator from the inner wall thereof to the outer wall thereof, thus connecting the interstice with a second chamber. Said second chamber is located within the cavity of the body of the pump, between the inner wall of the body of the pump and the outer wall of the stator, said second chamber being separated in a tightly sealed manner from the pressure chamber and being in communication with the suction inlet.

The groove extends along the inner wall of the stator and follows the direction of rotation of the rotor, from the point where the first duct exits from the inner wall to the point where the second duct starts from said inner wall. Said groove preferably has a decreasing thickness following said direction of rotation.

Preferably, the second duct starts from a point on the inner wall of the stator along a direction that is substantially tangent to said inner wall at said point. In addition, preferably, the second duct has a diameter smaller than the diameter of the first duct.

Thanks to this decreasing thickness of the groove and to the fact that the diameter of the second duct is smaller than the diameter of the first duct, the flow of pressurized fluid passing through the groove from the first duct to the second duct, by virtue of an effect similar to that of known hydraulic turbines, tends to bring into rotation the ring, thus making the latter rotate as far as possible at the same speed as the vanes, so as to limit the friction-related wear between vanes and ring and to achieve a consequent improvement in mechanical performance.

Preferably, the pump further comprises, on the outer surface of the ring of the pump, a series of protuberances, preferably arranged circumferentially at the height at which the first duct opens into said interstice. The pressurized fluid coming from the first duct engages on said protuberances, which form small “hydraulic vanes” and preferably have a size, in an axial direction, approximately corresponding to the size of said first duct. In this way, bringing into rotation of the rings is promoted. Said protuberances are, for example, teeth or the protuberances of a knurling.

Brief Description of Drawings

These and other features and advantages of the present invention will become evident from the following description of preferred embodiments given by way of non-limiting examples with reference to the annexed drawings, in which parts identified with identical or similar reference numerals denote parts having identical or similar function and construction, and in which:

Fig. 1 shows a plan view, partially in section, of a pump according to a first embodiment of the invention, in a state of maximum displacement;

Fig. 2 shows a plan view of the pump of Figure 1, in which, for clarity reasons, the stator, the rotating ring and the rotor with vanes have been shown in phantom lines by means of a dashed line;

Fig. 3 shows a plan view, partially in section, of a pump according to a first embodiment of the invention, in a state of minimum displacement;

Fig. 4 shows a plan view of the pump of Figure 3, in which, for clarity reasons, the stator, the rotating ring and the rotor with vanes have been shown in phantom lines by means of a dashed line;

Fig. 5 shows a plan sectional view of the stator of the pump according to the first embodiment of the invention;

Fig. 6 shows a perspective view of the stator of the pump according to the first embodiment of the invention;

Fig. 7 shows a plan view of a pump according to a second embodiment of the invention, without lid, in a state of maximum displacement;

Fig. 8 shows a plan view of the pump of Figure 6, without lid, in a state of minimum displacement;

Figs.9a-9b show perspective views of the rotating ring of the pump according to two different embodiments of the invention.

Description of Embodiments

A volumetric rotary pump 10 according to a first embodiment of the present invention is described below with reference to Figures 1-4.

The pump 10 is a volumetric rotary pump with variable displacement, configured, in a known manner, to pump a fluid, for example, lubricating oil, from a suction duct 11 to a discharge duct 12.

The pump 10 comprises a stator 20 housed in a cavity 14 formed in a body 13 of the pump 10. The stator 20 has an inner wall 21 defining, together with a bottom 15 of the body 13 of the pump 10 and a lid (not shown) of the pump, a preferably cylindrical pumping chamber 30. A vane rotor 40 is rotatably mounted within the pumping chamber 30 and eccentric thereto and is configured to rotate about an axis of rotation O thereof by being driven by a driving shaft (not shown), for example in a clockwise direction, as indicated by arrow R in Figures 1-4.

The suction duct 11 and the discharge duct 12, preferably provided in the body 13 of the pump 10, are connected to the pumping chamber 30 of the pump 10. In particular, the suction duct 11 and the discharge duct 12 are hydraulically connected to the pumping chamber 30 through a suction inlet 31 and a discharge outlet 32 of the pumping chamber 30, respectively. The suction inlet 31 and the discharge outlet 32 are provided on the bottom 15 of the body 13 of the pump 10, which bottom delimits the pumping chamber 30.

According to the invention, a ring 50 is also rotatably mounted within the pumping chamber 30. The ring 50 has an outer surface 51, facing the inner wall 21 of the stator 20, and an inner surface 52, facing the rotor 40.

The vane rotor 40 comprises, in a known manner, a plurality of slots 41, typically radial slots, open towards the outside of the rotor 40 and having a blind bottom 42 proximal to the axis of rotation O of the rotor 40. Each of the slots 41 is adapted to slidably accommodate a respective vane 45 of the rotor 40. Each vane 45 has a first end 45a, facing outwards and distal to the rotor 40, said vane being adapted, in use, to contact, in a tightly sealed manner, the inner surface 52 of the ring 50, in order to allow, in a known manner, transfer of the fluid to be pumped from the suction duct 11 to the discharge duct 12. A second end 45b of each vane 45 is, instead, facing towards the blind bottom 42 of the respective slot 41.

Advantageously, the effect of the friction between the first ends 45a of the vanes 45 and the inner surface 52 of the ring 50, makes the ring 50 rotate in the same direction of rotation R as the rotor 40. Accordingly, the relative sliding between the vanes 45 and the ring 50 is reduced, if not eliminated, thereby reducing the friction- related wear between vanes 45 and ring 50.

Preferably, the ring 50 and the vanes 45 are made of metal, such as steel. In this way, the vanes 45 and the ring 50 exhibit greater resistance to wear. The stator 20 is advantageously made of a material other, lighter and cheaper, than that of the vanes 45, preferably a thermoset resin (e.g., a phenolic resin), possibly enriched with mineral fibers or carbon fibers in order to reduce the coefficient of thermal expansion of said resin, which is known to be higher than that of metals, or the stator is made of graphite or carbon-graphite, possibly impregnated with other materials. In addition, the stator 20, not being constrained to the rotating ring 50, can be dimensioned, axially and radially, in such a way as to achieve minimum axial clearance and maximum radial clearance, by virtue of the coefficient of thermal expansion, at high temperatures (at which the pump usually operates), thus preserving free movement between the stator 20, the ring 50 and the cavity 14 of the body 13 of the pump 10, at the lowest temperatures of use, thus optimizing the volumetric efficiency of the pump 10 in the thermal field of operation.

In addition, the rotor 40 is associated with centering rings 46, mounted at the axial ends of the rotor 40 and configured, in a known manner, to contact the second ends 45b of the vanes 45, so as to keep the first ends 45a of the vanes 45 at a minimum distance from the inner surface 52 of the ring 50. In particular, the centering rings 46 promote, in a condition of low speed of rotation of the rotor 40, the approaching of the vanes 45 towards the ring 50, thus enabling the first triggering of the function of suction/compression of the fluid performed in a known manner by the pump 10.

In order to vary the displacement of the pump 10, the mutual position between the axis of rotation O of the rotor 40 and a central axis O’ of the pumping chamber 30 is changed. In particular, the axis of rotation O of the rotor 40 is preferably kept stationary relative to the body 13 of the pump 10, whereas the stator 20, and the ring 50 with it, are movably mounted in the cavity 14 of the body 13 of the pump 10, so as to cause movement of the pumping chamber 30 from a position in which the central axis O’ of the pumping chamber 30 is at a maximum distance from the rotation axis O of the rotor 40 and the rotor 40 is substantially tangent to the inner surface 52 of the ring 50 (maximum displacement), to a position in which the central axis O’ of the pumping chamber 30 is substantially coaxial with the axis of rotation O of the rotor 40 (minimum displacement).

In the embodiment shown in Figures 1-4, the stator 20 is mounted pivotally around a pivot axis W, external to the pumping chamber 30, and the variation in displacement is obtained by means of the pivoting of the stator 20 around said pivot axis W.

The pump 10 further comprises adjustment means to selectively control the eccentricity between the rotor 40 and the pumping chamber 30 and, therefore, the displacement of the pump.

In the embodiment shown in Figures 1-4, said adjustment means comprise a pressure chamber 60. Said pressure chamber 60 is a tightly sealed chamber located within the cavity 14 of the pump body 13, between an inner wall 16 of the pump body 13 and an outer wall 22 of the stator 20, and is delimited, following the direction of rotation R, between a seal 61, adapted to ensure tight sealing between the stator 20 and the inner wall 16 of the pump body 13, and a region 24 in which the stator 20 is pivotally joined or hinged to the pump body 13, so that said pressure chamber is in communication with the discharge outlet 32 and not with the suction inlet 21. Said pressure chamber 60 is therefore suitable for containing a pressurized fluid, produced by the pump 10 and suitable for exerting a thrust onto the stator 20 to make it rotate around the pivot axis W thereof. The adjustment means further comprise a spring 62 counteracting said thrust. The spring 62 tends to keep the stator 20 in the position in which the displacement is at its maximum (Figures 1 and 2), whereas the pressurized fluid pushes the stator 20 towards the position in which the displacement is at its minimum (Figures 3 and 4).

More specifically, the pressure of the fluid exiting the pump 10 increases according to the rotation speed and the permeability of a user’s circuit (not shown), fed by the pump 10, until the fluid, acting onto the outer wall 22 of the stator that delimits the pressure chamber 60 and overcoming the pre-load of the spring 62, causes the pivoting movement of the stator 20. Since the thrust generated by the pressurized fluid in the pressure chamber 60 begins to move the stator 20, the pressure increases proportionally with the rotation speed of the rotor 40 and the rigidity of the spring 62.

According to the invention, the pump 10 comprises a first duct 70, formed in the stator 20 and extending through the stator 20 from the outer wall 22 thereof to the inner wall 21 thereof, thus connecting the pressure chamber 60 to a narrow annular interstice 71 provided between the inner wall 21 of the stator 20 and the outer surface 51 of the ring 50. The first duct 70 allows the passage of pressurized fluid from the pressure chamber 60 to the aforesaid interstice 71, thus creating a bearing means between the inner wall 21 of the stator 20 and the outer surface 51 of the ring 50. This meatus makes it possible to reduce the friction between the inner wall 21 of the stator 20 and the outer surface 51 of the ring 50.

According to the embodiment shown in Figures 1-4, the first duct 70 exits from a point on the inner wall 21 of the stator 20 along a direction that is substantially tangent to said inner wall 21 at this exit point and in such a way that a flow of pressurized fluid reaching the interstice 71 through the first duct 70 has the same direction of rotation R as the rotor 40. In other words, the first duct 70 opens into the interstice 71 along a direction that forms an angle of substantially 0° with the direction tangent to the inner wall 21 at the aforesaid exit point of the first duct 70.

Advantageously, thanks to the angle at which the first duct 70 opens into the interstice 71, the pressurized fluid reaching the interstice 71, by acting onto the outer surface 51 of the ring 50, imparts a tangential thrust to said ring 50 in the same direction of rotation R in which the ring rotates, thereby promoting such rotation, which is useful especially at low rotational speeds of the rotor 40 or at low operational pressures of the pump. The rotation of the ring 50 determines, as already mentioned, less sliding of the first ends 45a of the vanes 45 over the inner surface 52 of the ring 50, and therefore less wear; this, in turn, results in higher performance of the pump 19 and lower energy consumption.

According to a variant, not shown, of the first embodiment, the first duct opens into the interstice along a direction that forms, with the direction tangent to the inner wall of the stator at the point of the inner wall where the first duct exits, an angle greater than 0° and smaller than 90°. In this variant, the tangential thrust imparted to the ring by the pressurized fluid is, again, in the same direction as the direction of rotation R of the ring, though it is smaller, at the same fluid pressure, than the thrust generated in the embodiment illustrated above, in which the first duct exits tangentially from the inner wall of the rotor.

In order to prevent the interstice 71 between the inner wall 21 of the stator 20 and the outer surface 51 of the ring 50 from causing a hydraulic short-circuit between the suction inlet 31 and the discharge outlet 32, said suction inlet 31 and discharge outlet 32, the stator 20 and the ring 50 are arranged and dimensioned so that said interstice 71 never simultaneously overlaps axially the suction inlet 31 and the discharge outlet 32, as shown in Figures 2 and 4 for the two boundary configurations of maximum and minimum displacement.

Referring particularly to Figures 5 and 6, the stator 20 further comprises, preferably, a second duct 72 and a groove 73 formed on the inner wall 21 of the stator 20.

The second duct 72, distanced from the first duct 70 by an arc of approximately 180°, passes through the stator 20 from the inner wall 21 thereof to the outer wall 22 thereof, thus connecting the interstice 71 with a second chamber 63. Said second chamber 63 is located within the cavity 14 of the pump body 13, between the inner wall 16 of the body 13 of the pump and the outer wall 22 of the stator 20, and is delimited, following the direction of rotation R, between the region 24 in which the stator 20 is pivotally joined, or hinged, to the pump body 13 and the seal 61, so that the second chamber 63 (and therefore the second duct 72) is in communication with the suction inlet 31 but not with the discharge outlet 32.

Preferably, the second duct 72 starts from a point on the inner wall 21 of the stator

20 along a direction that is substantially tangent to said inner wall 21 at said point. In addition, the second duct 72 preferably has a diameter smaller than the diameter of the first duct 70.

The groove 73 extends along the inner wall 21 of the stator 20 and follows the direction of rotation R, from the point where the first duct 70 exits from the inner wall

21 to the point where the second duct 72 starts from said inner wall 21. In addition, the groove 73 has a decreasing thickness following the direction of rotation R, i.e., it has a greater thickness at the first duct 70 and a smaller thickness at the second duct 72.

Thanks to the decreasing thickness of the groove 73 as well as to the fact that the diameter of the second duct 72 is smaller than the diameter of the first duct 70, the flow of pressurized fluid passing along the groove 73 from the first duct 70 to the second duct 72, by virtue of an effect similar to the one observed for the hydraulic turbines of known type, tends to bring the ring 50 into rotation, thus making said ring rotate as much as possible with a speed equal to the speed of the vanes 45, so as to limit the friction-related wear between vanes 45 and ring 50. According to an embodiment variant, not shown, the second duct starts from the inner wall of the stator along a direction that forms, with the direction tangent to the inner wall at the point where the second duct starts, an angle greater than 0° and smaller than 90°.

Referring to Figures 7 and 8, a further embodiment of a volumetric rotary vane pump with variable displacement according to the invention is described below.

The pump according to this further embodiment is indicated by reference numeral 110 and, correspondingly, the pump elements 110 already mentioned in respect of the pump 10 are indicated by reference numerals similar to the ones already used.

The pump 110 comprises a stator 120, housed in a cavity 114 formed in a body 113 of the pump 110, and a stator 120. The stator 120 has an inner wall 121 defining, together with a bottom 115 of the body 113 of the pump 110 and a pump lid (not shown), a preferably cylindrical pumping chamber 130. A vane rotor 140 is rotatably mounted within the pumping chamber 130 and eccentric thereto and is configured to rotate about an axis of rotation P thereof by being driven by a driving shaft (not shown), for example in a counterclockwise direction, as indicated by arrow S in Figure 7.

A suction duct and a discharge duct, both not shown, are hydraulically connected to the pumping chamber 130 of the pump 110 by means of a suction inlet 131 and a discharge outlet 132 of the pumping chamber 130, respectively. The suction inlet 131 and the discharge outlet 132 are provided on the bottom 115 of the pump body 113 delimiting the pumping chamber 130.

According to the invention, a ring 150 is also rotatably mounted within the pumping chamber 130. The ring 150 has an outer surface 151, facing the inner wall 121 of the stator 120, and an inner surface 152, facing the rotor 140.

The vane rotor 140 is similar to the rotor 40 of the first embodiment. Therefore, it comprises a plurality of slots 141 with blind bottoms 142, a plurality of vanes 145 having first ends 145a adapted, in use, to contact, in a tightly sealed manner, the inner surface 152 of the ring 150, and second ends 145b facing towards the blind bottom 142 of the respective slots 141. In addition, in this embodiment, too, the rotor 140 is associated with centering rings 146. For a more detailed description of the aforementioned elements, reference is made to what was described above for the first embodiment. As in the first embodiment, the ring 150 and the vanes 145 are preferably made of the same material, preferably metal, for example, steel, which provides greater wear resistance, whereas the stator 120 is made of a lighter and cheaper material, preferably a thermoset resin (for example, a phenolic resin), which is possibly enriched with mineral or carbon fibers to further reduce its thermal expansion coefficient, which is known to be higher than that of metals, or the stator is made of graphite or carbon- graphite, possibly impregnated with other materials. In addition, the stator 120, not being constrained to the rotating ring 150, can be dimensioned, axially and radially, in such a way as to achieve minimum axial clearance and maximum radial clearance, by virtue of the coefficient of thermal expansion, at high temperatures (at which the pump usually operates), thus preserving free movement between the stator 120, the ring 150 and the cavity 114 of the body 113 of the pump 110, at the lowest temperatures of use, thus optimizing the volumetric efficiency of the pump 110 in the thermal field of operation.

In this embodiment too, advantageously, the effect of the friction between the first ends 145a of the vanes 145 and the inner surface 152 of the ring 150 causes the ring 150 to come into rotation, in the same direction of rotation S as the rotor 140. Accordingly, the relative sliding between the vanes 145 and the ring 150 is reduced, if not eliminated, thereby reducing the friction-related wear between vanes 145 and ring 150.

In order to vary the displacement of the pump 110, the mutual position between the axis of rotation P of the rotor 140 and a central axis P’ of the pumping chamber 130 is changed. As in the previous embodiment, the axis of rotation P of the rotor 140 is preferably kept stationary relative to the body 113 of the pump 110, whereas the stator 120, and the ring 150 with it, are movably mounted in the cavity 114 of the body 113 of the pump 110, so as to cause movement of the pumping chamber 130 from a position in which the central axis P’ of the pumping chamber 130 is at a maximum distance from the rotation axis P of the rotor 140 and the rotor 140 is substantially tangent to the inner surface 152 of the ring 150 (maximum displacement) to a position in which the central axis P’ of the pumping chamber 130 is substantially coaxial with the axis of rotation P of the rotor 140 (minimum displacement).

In the embodiment illustrated in Figures 7 and 8, the stator 120 is translationally mounted and the variation in displacement is obtained by means of the translation of the stator 120.

The pump 110, too, comprises adjustment means to selectively control the eccentricity between the rotor 140 and the pumping chamber 130 and, therefore, the displacement of the pump. Similarly to what was described above for the first embodiment shown, said adjustment means of the pump 110 comprise a pressure chamber 160, located within the cavity 114 of the body 113 of the pump, between an inner wall 116 of the body 113 of the pump and an outer wall 122 of the stator 120. Said pressure chamber 160 is a tightly sealed chamber suitable for containing a pressurized fluid, produced by the pump 110 and suitable for exerting a thrust onto the stator 120 to make it translate. The pump 110 further comprises a spring 162 counteracting said thrust. The spring 162 tends to keep the stator 120 in the position in which the displacement is at its maximum (Figure 7), whereas the pressurized fluid pushes the stator 120 towards the position in which the displacement is at its minimum (Figure 8).

The pressure of the fluid exiting the pump 110 depends on the speed of rotation of the rotor 140 as well as on the rigidity of the spring 162, as shown for the first embodiment.

According to the invention, the pump 110 comprises a first duct 170 (shown only in Figure 8), formed in the stator 120 and extending through the stator 120 from the outer wall 122 thereof to the inner wall 121 thereof, thus connecting the pressure chamber 160 to a narrow annular interstice 171 provided between the inner wall 121 of the stator 120 and the outer surface 151 of the ring 150. The first duct 170 allows the passage of pressurized fluid from the pressure chamber 160 to the aforesaid interstice 171, thus creating a bearing means between the inner wall 121 of the stator 120 and the outer surface 151of the ring 150.

According to this embodiment, the first duct 170 opens into the interstice 171 and exits from a point on the inner wall 121 of the stator 120 along a direction substantially tangent to said inner wall 121 at this exit point and in such a way that a flow of pressurized fluid reaching the interstice 171 through the first duct 170 has the same direction of rotation S as the rotor 140. In other words, the first duct 170 opens into the interstice 171 along a direction that forms an angle of substantially 0° with the direction tangent to the inner wall 121 at the aforesaid exit point of the first duct 170.

According to a variant, not shown, of this embodiment, the first duct opens into the interstice along a direction that forms, with the direction tangent to the inner wall of the stator at the point of the inner wall where the first duct exits, an angle greater than 0° and smaller than 90°.

The advantageous effects of the first duct 170 are the same as those described for the first embodiment. The presence of fluid in the interstice 171 between the inner wall 121 of the stator 120 and the outer surface 151 of the ring 150 makes it possible to reduce the friction generated by the rotation of the ring 150 relative to the stator 120. In addition, thanks to the angle at which the first duct 170 opens into the interstice 171, the pressurized fluid reaching the interstice 171 imparts a tangential thrust to the ring 150 in the same direction of rotation S in which the ring rotates, thereby promoting such rotation and reducing the sliding of the first ends 145a of the vanes 145 over the inner surface 152 of the ring 150 and the wear resulting therefrom.

As with the first embodiment, in order to prevent the interstice 171 between the inner wall 121 of the stator 120 and the outer surface 151 of the ring 150 from causing a hydraulic short-circuit between the suction inlet 131 and the discharge outlet 132, said suction inlet 131 and discharge outlet 132, the stator 120 and the ring 150 are arranged and dimensioned so that said interstice 171 never simultaneously overlaps axially the suction inlet 131 and the discharge outlet 132, as shown in Figures 7 and 8 for the two boundary configurations of maximum and minimum displacement.

Similarly to what was described above for the first embodiment shown, the stator 120 of the pump 110, too, preferably comprises a second duct 172 and a groove (not shown) that is formed on the inner wall 121 of the stator 120.

The second duct 172, distanced from the first duct 170 by an arc of approximately 180°, extends through the stator 120 from the inner wall 121 thereof to the outer wall 122 thereof, thus connecting the interstice 171 to a second chamber 163. Said second chamber 163 is located within the cavity 114 of the body 113 of the pump, between the inner wall 116 of the pump body 113 and the outer wall 122 of the stator 120, between the inner wall 116 of the pump body 113 and the outer wall 122 of the stator 120, and is separated, in a tightly sealed manner, from the pressure chamber 160, so that the second chamber 163 is in communication with the suction inlet 131 but not with the discharge outlet 132.

Preferably, the second duct 172 starts from a point on the inner wall 121 of the stator 120 along a direction that is substantially tangent to said inner wall 121 at said point. In addition, the second duct 172 preferably has a diameter smaller than the diameter of the first duct 170.

The groove extends along the inner wall 121 of the stator 120 and follows the direction of rotation S, from the point where the first duct 170 exits from the inner wall 121 to the point where the second duct 172 starts from said inner wall 121. In addition, the groove has a decreasing thickness following the direction of rotation S, i.e., it has a greater thickness at the first duct 170 and a smaller thickness at the second duct 172.

Thanks to the decreasing thickness of the groove as well as to the fact that the diameter of the second duct 172 is smaller than the diameter of the first duct 170, the flow of pressurized fluid passing along the groove from the first duct 170 to the second duct 172, by virtue of an effect similar to the one observed for the hydraulic turbines of known type, tends to bring the ring 150 into rotation, thus making said ring rotate as much as possible with a speed equal to the speed of the vanes 145, so as to limit wear between vanes 145 and ring 150.

According to an embodiment variant, not shown, the second duct starts from the inner wall of the stator along a direction that forms, with the direction tangent to the inner wall at the point where the second duct starts, an angle greater than 0° and smaller than 90°.

Referring to Figure 9a, preferably, both in the first and in the further embodiment shown, it is possible to provide, on the outer surface 51 (or 151) of the ring 50 (or 150), a series of teeth 53, forming small “hydraulic vanes”, on which the pressurized fluid coming from the first duct 70 (or 170) engages, thereby promoting the bringing into rotation of the ring 50 (or 150). This is of particular advantage at low rotational speeds of the rotor 40 (or 140), or at low operational pressures, when it is more difficult to bring into rotation the ring 50 (or 150). For example, this series of teeth 53 runs along a circumferential groove 54 formed on said outer surface 51 (or 151) of the ring 50 (or 150), so that said teeth 53 do not protrude radially beyond the outer surface 51 (or 151) of the ring, thus preventing possible cogging or bouncing between ring 50 (or 151) and stator 20 (or 120).

Referring to Figure 9b, as an alternative to the series of teeth 53, on the outer surface 51 (or 151) of the ring 50 (or 150) there is provided a knurled strip 55, the protuberances of which act as small “hydraulic vanes” on which the pressurized fluid coming from the first duct 70 (or 170) engages, thereby promoting the bringing into rotation of ring 50 (or 150). In this case, too, the knurled strip 55 runs along a circumferential groove 56 formed on said outer surface 51 (or 151) of the ring 50 (or 150), so that these protuberances of the knurled strip 55 do not protrude radially beyond the outer surface 51 (or 151) of the ring, thus preventing possible cogging or bouncing between ring 50 (or 151) and stator 20 (or 120).

This series of teeth 53 or this knurled strip 55 is preferably located at the height at which the first duct 70 (or 170) opens into the interstice 71 (or 171), in order to maximize the thrust exerted by the pressurized fluid coming from the first duct 70 (or 170).

Therefore, in the embodiments in which both the teeth 53, or the knurled strip 55, and the groove 73 are provided between the first duct 70 (or 170) and the second duct 72 (or 172), said groove 73 faces the circumferential groove 54 along which said series of teeth 53 runs or the knurled strip 55.