Patent Grant
9404495
This application is a National Stage of International Application No. PCT/IB2013/051977 filed Mar. 13, 2013, claiming priority based on Italian Patent Application Nos. TO2012A000236, filed Mar. 19, 2012 and TO2012A001007, filed Nov. 20, 2012, the contents of all of which are incorporated herein by reference in their entirety.
The present invention relates to variable displacement pumps, and more particularly it concerns a rotary positive displacement pump of the kind in which the displacement variation is obtained by means of the rotation of an eccentric ring (stator ring).
Preferably, but not exclusively, the present invention is employed in a pump for the lubrication oil of a motor vehicle engine.
It is known that, in pumps for making lubricating oil under pressure circulate in motor vehicle engines, the capacity, and hence the oil delivery rate, depends on the rotation speed of the engine. Hence, the pumps are designed so as to provide a sufficient delivery rate at low speeds, in order to ensure lubrication also under such conditions. If the pump has fixed geometry, at high rotation speed the delivery rate exceeds the necessary rate, whereby a high power absorption, and consequently a higher fuel consumption, and a greater stress of the components due to the high pressures generated in the circuit occur.
In order to obviate this drawback, it is known to provide the pumps with systems allowing a delivery rate regulation at the different operating conditions of the vehicle, in particular through a displacement regulation. Different solutions are known to this aim, which are specific for the particular kind of pumping elements (external or internal gears, vanes . . . ).
A system often used in rotary pumps employs a stator ring with an internal cavity, eccentric relative to the external surface, inside which the rotor, in particular a vane rotor, rotates, the rotor being eccentric with respect to the cavity under operating conditions of the pump. By rotating the stator ring by a given angle, the relative eccentricity between the rotor and the cavity, and hence the displacement, is made to vary between a maximum value and a minimum value, substantially tending to zero (stall operating condition). A suitably calibrated opposing resilient member allows the rotation when a predetermined delivery rate is attained and makes the pump substantially deliver such a predetermined delivery rate under steady state conditions. A pump of this kind is disclosed for instance in U.S. Pat. No. 2,685,842.
U.S. Pat. No. 4,406,599 discloses a pump with a pair of stator rings arranged side by side and having respective oval cavities, which are mutually aligned in a maximum displacement condition of the pump. The displacement is made to vary by rotating the rings relative to each other in opposite directions by means of gears or racks, external to the pump, which mesh with teeth formed on the external surfaces of the rings. The rotation is driven by a piston responsive to the pressure conditions in a circuit utilising the pumped fluid.
The presence of external control members makes such a prior art pump complex and relatively cumbersome.
It is an object of the present invention to provide a variable displacement pump with double eccentric ring, and a method of regulating the displacement of such a pump, which obviate the drawbacks of the prior art.
According to the invention, this is obtained in that the stator ring is housed within an eccentric cavity of an external ring, which is configured as a multistage rotary piston for displacement regulation, arranged to be directly driven by a fluid under pressure in order to be rotated within a predetermined angular interval and arranged to transmit the rotary motion to the stator ring in order to make it rotate in opposite direction to the external ring.
Advantageously, at least one piston stage may have an actuating surface, onto which the fluid under pressure acts, having an area which changes during the piston rotation.
Preferably, for the transmission of the rotation to the stator ring, facing surfaces of the external ring and the stator ring have formed thereon respective toothed sectors with which an idle toothed wheel meshes, the toothed sector of the external ring being concentric with the external surface of the ring and the toothed sector of the stator ring being formed on an arc of an involute resulting from a composition of the relative rotations of the eccentricities of the cavities of both rings.
The rotation of the external ring is opposed by a flat spiral spring, which may be a bimetallic spring so as to exhibit a temperature-dependent behaviour.
The invention also implements a method of regulating the displacement of a rotary positive displacement pump by means of the rotation of an eccentric stator ring inside which the rotor rotates, the method comprising the steps of:
Advantageously, the step of directly controlling the piston rotation by means of fluid under pressure includes at least:
According to a further aspect of the invention, there is also provided a lubrication system for a motor vehicle engine, in which the adjustable displacement pump and the method of regulating the displacement set forth above are employed.
Further features and advantages of the invention will become apparent from the following description of preferred embodiments, given by way of non limiting examples with reference to the accompanying drawings, in which:
Referring to
In the present description, the term “coaxial or substantially coaxial” is used to denote a minimum distance, tending to O, between centres O and O′.
Advantageously, eccentric rings 12 and 112 are mounted in such a manner that, in the minimum displacement position shown in
Rotor 15 has a set of vanes 16, radially slidable in respective radial slots. At an outer end, vanes 16 are at a minimum distance from side surface 113a of cavity 113, whereas at the inner end they rest on guiding or centring rings 17, mounted at the axial ends of rotor 15 and arranged to maintain the minimum distance between vanes 16 and surface 113a under any condition of eccentricity. Also centring rings 17 will be coaxial or substantially coaxial with rotor 15 in the minimum displacement position.
A suction chamber 18, communicating with a suction duct 20, and a delivery chamber 19, communicating with a delivery duct 21, are defined between rotor 15 and surface 113a. Such chambers are substantially symmetrical and have phasings that are ideal for the maximum volumetric efficiency, as it is clearly apparent for the skilled in the art.
Rings 12 and 112, as well as centring rings 17 and rotor 15, are preferably formed by a process of metal powder sintering, or by moulding thermoplastic or thermosetting materials, with possible suitable finishing operations on some functional parts, according to the dictates of the art.
In order to control the rotation of external ring 12, the latter has on its external surface a pair of radial appendages 23, 24, which project into respective chambers 25, 26 defined by ring 12 and by respective recesses in the side surface of cavity 11 and slide onto bases 25a, 26a of chambers 25, 26, respectively. Such appendages may be integral parts of ring 12 or they may be separate elements, fastened to the ring, or yet radially slidable vanes, which are guided in suitable radial slots formed in ring 12 and are suitably pushed into contact with bases 25a, 26a of chambers 25, 26 by resilient means. In the region where they are in contact with the base of the respective chamber, appendages 23, 24 may be equipped with gaskets 27, 28, respectively, for optimising the hydraulic seal.
One of the chambers (in the illustrated example, chamber 25) is permanently connected to delivery chamber 19, through a duct 50, or preferably to the members utilising the pumped fluid (in particular, in the preferred application, to a point of the lubrication system located downstream the oil filter), through a first regulation duct, not shown in these Figures, ending into an inlet passage 29. By means of a valve operated by the electronic control unit of the vehicle, the other chamber can in turn be put in communication with the members utilising the pumped fluid, through a second regulation duct ending into an inlet passage 30. Also the valve and the second regulation duct are not shown in these Figures.
Both appendages 23, 24 are therefore exposed to the fluid pressure conditions existing at the delivery side and/or in the utilisation members and they form a first and a second stage of displacement regulation, respectively, the second stage operating jointly with the first stage, as it will be better explained in the description of the operation. The radial size and the circumferential amplitudes of chambers 25, 26 will be determined by the operation characteristics required from the pump. Chambers 25, 26 can also be defined as regulation cylinders, and appendages 23, 24 form the corresponding pistons. One appendage (appendage 23 in the drawing) may be provided with projections 23a, 23b acting as stops in the rest position and in the operating condition, respectively, and keeping the appendage spaced apart from the adjacent end wall of chamber 25 at the end of the ring stroke.
Both chambers 25, 26 are equipped with drainage ducts 31, 32 for discharging oil seepages, if any, and for compensating volume variation generated when ring 12 is made to rotate.
In the illustrated embodiment, drains 31, 32 communicate with the outside of the pump. In other embodiments, drains 31, 32 are for instance connected to the suction chamber.
If necessary, means are provided for adjusting the drainage flows in order to damp possible hydraulic pulsations of the displacement regulating system.
Toothed sectors 51, 52 are formed on facing surfaces of rings 12, 112 and an idle toothed wheel 53 is interposed between said sectors. The “driving” toothed sector 51 is concentric with the external surface of ring 12, guided within chamber 11, whereas the “driven” toothed sector 52 is formed on the arc of the involute resulting from the composition of the relative rotations of the eccentricities of cavities 13, 113. If the eccentricities are the same, during the relative rotation of the rings centre O′ of cavity 112 will then move along a rectilinear trajectory.
Referring to
It is to be appreciated that, during the regulation rotation, spiral spring 34, thanks to the negligible variation of the twisting torque and to the transmission ratio of the gear mechanism, will undergo negligible variations of its torque opposing the hydraulic torque of the rotary piston.
Advantageously, spring 34 may be made of a bimetallic material, so that its characteristic may suitably change depending on the operation temperature.
Turning to
It is pointed out that the choice of connecting chamber 25 directly to delivery duct 21 or, in the alternative, to outlet 63 of the oil filter depends on the requirements defined by the engine manufacturer. However, the connection to the filter outlet is the choice ensuring the greatest stability in the regulation pressure since, as known, due to the nature of the positive displacement pumps, the delivery pressure has surges which are damped by filter 62. Moreover, as a skilled in the art will readily appreciate, the displacement regulation is independent of any pressure drop caused by the filter, for instance due to the greater or smaller clogging thereof because of impurities, or due to changes in oil viscosity.
Moreover, valve 66 might be housed in the body of pump 1, in which case ducts 64, 65 will be passages formed in said body.
The operation of pump 1 is as follows.
Under rest conditions, pump 1 is in the condition shown in
The delivery pressure or the pressure downstream oil filter 62 are brought to chamber 25 through duct 50 or 64 and they will act on appendage 23, thereby creating an hydraulic thrust on ring 12 and generating a rotation torque. Once the calibration value of the counteracting spring 34 has been attained, such a torque will cause a rotation of ring 12, in this case in clockwise direction, which rotation will be transmitted to ring 112 through idle wheel 53 meshing with toothed sectors 51 and 52 and will make ring 112 rotate in counterclockwise direction by the same angle. If, as it has been assumed, the eccentricities of cavities 13 and 113 relative to the external surfaces of the respective rings are the same, the rotation of ring 112 will cause a rectilinear translation of centre O′ towards the right, proportional to the amount of the rotation, thereby proportionally reducing the distance between rotor 15 and cavity 113 and consequently the pump displacement, and stabilising the pressure at the calibration value. As parameters such as the speed, the fluidity/temperature of the fluid, the engine “permeability” (intended as the amount of oil used by the engine) and so on change, such a pressure will be maintained and controlled through the variation of the eccentricity and hence of the displacement.
When, as a function of the different operating parameters of the engine, as detected by the electronic control unit of the vehicle, it is desired to operate at a lower pressure value, with a consequent reduction in the absorbed power, fluid under pressure can be fed also to chamber 26 by means of valve 66, whereby a supplementary hydraulic thrust concordant with the thrust exerted on appendage 23 is created on appendage 24. In this way, the rotation torque of the piston is increased and the pump displacement is reduced. Stopping the feed to chamber 26 will bring the pressure back to the previous higher value through the variation of the displacement.
The rotation of the rings may continue until the position shown in
By mutually exchanging the drains and the oil inlets to chambers 25, 26, it is also possible to generate torques adding to the counteracting torque generated by spring 34.
An important parameter in managing the delivery rate/pressure of an oil pump for thermal engines is temperature, the increase of which makes the oil become more fluid and the engine permeability increase. Consequently, the pump displacement should proportionally increase. This may be assisted if the opposing load of spring 34 increases. In order to obtain this, flat spiral spring 34 may be made of a bimetallic material such that temperature causes an increase in the rigidity and hence in the counteracting torque. In order to obtain the change in the rigidity, the small oil flow rate for the lubrication of shaft 54 of idle wheel 53 may be exploited: the oil, after having licked casing 33 of spring 34 and having transmitted its temperature to the same spring, can freely discharge to the suction chamber through drain 58.
In the pump described above, bases 25a, 26a of chambers 25, 26, when viewed in plan, are arcs of circumference the centre of which is located on the rotation axis of ring 12, and chambers 25, 26 have constant radial sizes. This entails that the different stages or pistons have actuating surfaces, on which the fluid under pressure acts, having constant areas and therefore generate a torque that is proportional to the pressure of the actuating fluid and is constant over the whole rotation of ring 12.
In the pump according to this embodiment, denoted 401, the displacement regulation pistons consist of radially slidable vanes 423, 424, which are guided in respective seats 423′, 424′ and are pushed into contact with bases 425a, 426a of chambers 425, 426 by resilient means 470, 471, for instance spiral or leaf springs. Bases 425a, 426a, when viewed in plan, are shaped as arcs of circumferences the centres of which do not coincide with the centre of rotation of ring 12, and therefore the chambers have variable radial sizes (in particular, in the Figure, radial sizes steadily increasing in the direction of the rotation performed by ring 12 for bringing the pump from the maximum displacement position to the minimum displacement position). The arcs forming bases 425a, 426a may possibly have different radiuses. It is also possible that only one chamber (in particular, the chamber in which the stage permanently exposed to the fluid pressure moves, for instance chamber 425) has a variable radial size. The skilled in the art will have no problem in designing and sizing vanes 423, 424 and resilient elements 470, 471 so as to ensure the contact between the vanes and bases 425a, 426a of chambers 425, 426 along the whole of the arc of rotation of ring 12.
It is to be appreciated that, in the illustrated example, one of the vanes (for instance vane 423) is inserted in radial appendage 23, whereas vane 424 is directly inserted in ring 12. In other embodiments, both vanes 423, 424 may be inserted in ring 12 or in the respective appendage 23, 24.
The operation of such a variant embodiment is similar to that described above. Considering vane 423, the difference is that, during rotation, due to the lack of concentricity of wall 425a with respect to ring 12 and hence to the increasing radial size of chamber 425, vane 423 will progressively come out from slot 423′, whereby its actuating area (and of course its thrust area) and consequently the rotation torque applied to ring 12 progressively increase. This allows compensating, for instance, the increase in the resistant torque caused by the increase in the force exerted by reaction spring 34 and/or by the rotation frictions. What has been stated for vane 423 applies of course also to vane 424.
The invention actually attains the desired aims. By configuring external ring 12 as a multistage rotary piston to which the pressure of the control fluid is directly applied, and by driving stator ring 112 by means of external ring 12, external driving units are eliminated, and hence the structure is simpler and therefore less expensive and less prone to failures, as well as less cumbersome. Both rings, with substantially circular cross section, may be made with limited radial thicknesses. A further limitation in the radial overall size is obtained by configuring the rings so that the movement of the axis of centre O′ takes place on a rectilinear trajectory.
It is clear that the above description has been given only by way of non-limiting example and that changes and modifications are possible without departing from the scope of the invention.
For instance, even if in the illustrated embodiment shaft 15a of rotor 15 is guided by body 10 whereas spiral spring 34 with the calibration means consisting of casing 33 and ring nut 55 are housed within cover 14, the arrangement could be reversed, or also the spring and the calibration means could be housed within body 10.
Moreover, body 10 might be a through element, which could be possibly formed by means of extrusion or moulding technologies, and might be closed at its ends by suitable covers, centred and aligned by suitable centring means, for instance pegs.
Furthermore, external ring 12 could have, in correspondence of appendages 23 and 24 (or vanes 423, 424), a lightening cavity housing a barrier rigidly connected to the body and communicating with one of chambers 25, 26 (or 425, 426) in order to receive the fluid under pressure fed to such a chamber, so as to offer a greater overall thrust surface. Such a lightening cavity, and possible further similar cavities formed at the periphery of ring 12, could be connected instead to the delivery side of the pump or to the outlet of the oil filter in order to form further regulations stages, preferably controlled from the outside in similar manner to the stage consisting of appendage 24 and chamber 26.
An inversion between the supply and the drains in at least one of the stages could also be possible, so as to add/subtract the actuating torques, thereby allowing the attainment of several variants for the pump calibration and management. Moreover, it is also possible to form radial chambers, steadily connected to the delivery duct under pressure, in order to counterbalance the radial hydraulic thrusts acting on the eccentric rings.
Moreover, even though
If, in the embodiment with adjustable thrust, lightening cavities with a barrier shaped so as to give rise to further regulation stages are provided, also such stages may have variable thrust areas.
Lastly, even if the invention has been disclosed in detail with reference to a pump for the lubrication oil of a motor vehicle engine, it may be applied to any positive displacement pump for conveying fluid from a first to a second working environment, in which a delivery rate reduction as the pump speed increases is convenient.
| Number | Date | Country | Kind |
|---|---|---|---|
| T02012A0236 | Mar 2012 | IT | national |
| TO2012A1007 | Nov 2012 | IT | national |
| Filing Document | Filing Date | Country | Kind |
|---|---|---|---|
| PCT/IB2013/051977 | 3/13/2013 | WO | 00 |
| Publishing Document | Publishing Date | Country | Kind |
|---|---|---|---|
| WO2013/140305 | 9/26/2013 | WO | A |
| Number | Name | Date | Kind |
|---|---|---|---|
| 2685842 | Hufferd | Aug 1954 | A |
| 3234816 | Von Thuengen | Feb 1966 | A |
| 3334546 | Vuolle-Apiala | Aug 1967 | A |
| 4177024 | Lohn | Dec 1979 | A |
| 4406599 | Stephan | Sep 1983 | A |
| 4778352 | Nakajima | Oct 1988 | A |
| Number | Date | Country |
|---|---|---|
| 785548 | Dec 1980 | SU |
| Entry |
|---|
| International Search Report for PCT/IB2013/051977 dated Jun. 21, 2013. |
| Written Opinion of PCT/IB2013/051977 dated Jun. 21, 2013. |
| Number | Date | Country | |
|---|---|---|---|
| 20150050173 A1 | Feb 2015 | US |
