Patent Publication Number: US-9404495-B2

Title: Variable displacement pump with double eccentric ring and displacement regulation method

Description:
CROSS REFERENCE TO RELATED APPLICATIONS 
     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. 
     TECHNICAL FIELD 
     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. 
     PRIOR ART 
     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. 
     DESCRIPTION OF THE INVENTION 
     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:
         providing an external ring having an eccentric cavity within which the stator ring is housed;   configuring the external ring as a multistage rotary piston;   directly controlling the piston rotation by means of fluid under pressure; and   transmitting the rotation of the external ring to the stator ring in such a manner that the two rings rotate in opposite directions.       

     Advantageously, the step of directly controlling the piston rotation by means of fluid under pressure includes at least:
         applying the fluid to a first stage of the piston in order to maintain the displacement at a first value determined through a suitable calibration of members opposing the rotation; and   applying the fluid to a second stage of the piston, simultaneously with the application to the first stage and upon an external command, in order to bring the displacement to a second value different from the first one.       

     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. 
    
    
     
       BRIEF DESCRIPTION OF THE FIGURES 
       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: 
         FIG. 1  is a plan view of a pump according to the invention, from which the cover has been removed, in the maximum displacement condition; 
         FIG. 2  is a view similar to  FIG. 1 , in the minimum displacement condition; 
         FIG. 3  is a plan view, similar to  FIG. 2 , showing the displacement regulation mechanism integrated in the cover; 
         FIG. 4  is a cross-sectional view of the pump according to a plane passing through line Y-Y in  FIG. 3 ; 
         FIGS. 5 and 6  are diagrams of a lubrication circuit of a motor vehicle engine using the pump according to the invention, relative to the maximum and minimum displacement condition, respectively; and 
         FIGS. 7 and 8  are views similar to  FIGS. 1 and 2 , relating to a variant embodiment. 
     
    
    
     DESCRIPTION OF PREFERRED EMBODIMENTS 
     Referring to  FIGS. 1 and 2 , a pump according to the invention, generally denoted by reference numeral  1 , includes a body  10  having a cavity  11  with substantially circular cross-section in which a first movable ring  12  (external ring) is located, which in turn has an axial cavity  13 , also with substantially circular cross-section, eccentrically arranged relative to cavity  11 . A second movable ring  112  (stator ring) is located in cavity  13 , which ring in turn it has an axial cavity  113 , also with substantially circular cross-section, eccentrically arranged relative to cavity  13  and having a centre O′. Rings  12  and  112  are arranged to rotate in mutually opposite directions by a certain angle in order to vary the pump displacement, as it will be better disclosed below. In particular, ring  12  acts as a multistage rotary piston and is arranged to cause the rotation of internal ring  112 , acting as an eccentric stator ring. Cavity  113  in turn houses a rotor  15 , rigidly connected to a driving shaft  15   a  making it rotate about a centre O, for instance in clockwise direction, as shown by arrow F. In a maximum displacement position (shown in  FIG. 1 ), centres O and O′ are located on a same axis and are mutually spaced apart, and rotor  15  is substantially tangent to side surface  113   a  of cavity  113 . In a minimum displacement position (shown in  FIG. 2 ), rotor  15  and cavity  113  are coaxial or substantially coaxial. 
     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  FIG. 2 , external ring  12  is oriented so that its minimum radial thickness is located at the top in the Figure and internal ring  112  is oriented so that its minimum radial thickness is located at the bottom in the Figure. Otherwise stated, the eccentricities of the respective cavities  13 ,  113  are offset by 180°. Preferably, cavities  13 ,  113  have the same eccentricity relative to the external surface of the respective ring. 
     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  113   a  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  113   a  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  113   a . 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  25   a ,  26   a  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  25   a ,  26   a  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  23   a ,  23   b  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  FIGS. 3, 4 , idle wheel  53  cooperates with a member  34  opposing the rotation of ring  12 , in particular a flat spiral spring, preloaded so as to prevent the rotation of the ring as long as the pressure applied to appendage  23  (or the overall pressure applied to appendages  23  and  24 ) is lower than a predetermined threshold. Spiral spring  34  is located in a casing  33  that, in the illustrated exemplary embodiment, is fastened to a cover  14  closing one end of cavities  11 ,  13  and  113 , which, in the illustrated example, are blind cavities. The inner end portion of spring  34  is so shaped as to be coupled with the end portion of shaft  54  of idle wheel  53 , whereas the outer end portion is locked to the internal wall of casing  33 . The latter may be rotated, for instance by using a dynamometric key, in order to adjust the preloading of spring  34 . A ring nut  55  allows blocking casing  33  in the desired calibration position, independently of the constructive tolerances of the whole mechanism. A sealing gasket  56  is moreover provided between casing  33  and cover  14  in order to isolate internal chamber  57  of the same casing from the outside. A drain  58  puts such a chamber in communication with suction chamber  18 , for the aims that will be disclosed below. 
     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  FIGS. 5 and 6 , lubrication circuit  100  of a motor vehicle engine  60  using pump  1  is shown. Reference numerals  61  and  62  denote the oil sump and the oil filter, connected in conventional manner to suction and delivery ducts  20 ,  21  through ducts denoted by the same reference numerals, and reference numeral  63  denotes the outlet duct of filter  62 , conveying the oil to engine  60 . A first branch of outlet  63  of oil filter  62  forms the first regulation duct  64 , which conveys the oil to chamber  25  and can be used in the alternative to passage  50 . A second branch of outlet  63  of oil filter  62  forms the second regulation duct  64 , in which valve  66  controlled by the electronic control unit, for instance an electromagnetic valve, is connected. Depending on the position of such a valve, oil leaving filter  62  may be conveyed to chamber  26  or intercepted: in the latter case, the oil present in chamber  25  and in duct  65  may be sent back to oil sump  61  through valve  66  and duct  67 . 
     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  FIG. 1 . As said, centre of rotation O of rotor  15  is offset relative to centre O′ of cavity  113  of eccentric ring  112  and rotor  15  is located close to wall  113   a  of the cavity. When pump  1  is started, the clockwise rotation of rotor  15  will give rise to an oil flow through chamber  19  and the associated delivery duct  21  and, at the same time, an equal volume of oil will be sucked from chamber  18  and the associated suction duct  20 . As the rotation speed and the flow rate increase, the lubrication system of the engine, by opposing an increasing resistance to the flow, will make the pressure increase. 
     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  FIG. 2  is attained, where projection  23   b  of appendage  23  is in contact with the wall of chamber  25 , centres O and O′ coincide and vanes  16  and centring rings  17  rotate with the rotor without changes in their radial relative position. Consequently, the displacement is null and the pump is in stall condition. It is to be pointed out that this position may be taken when a hydraulic lock of the delivery pressure is approaching. In the constructional practice, a minimum displacement is preferably maintained by protecting the pump with a maximum pressure valve. 
     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  25   a ,  26   a  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 . 
       FIGS. 7 and 8  show an embodiment in which the torque applied to ring  12  may be changed during the displacement regulation in order to take into account possible changes in the resistant torques encountered during such a regulation, for instance due to changes in the resistance opposed by opposing spring  34  and/or in the rotation frictions. 
     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  425   a ,  426   a  of chambers  425 ,  426  by resilient means  470 ,  471 , for instance spiral or leaf springs. Bases  425   a ,  426   a , 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  425   a ,  426   a  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  425   a ,  426   a  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  425   a  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  15   a  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  FIGS. 7 and 8  show chambers  425 ,  426  with bases  425   a ,  426   a  consisting of arcs of circumferences arranged so that such chambers have progressively increasing radial sizes in the direction of rotation of ring  12  from the maximum displacement position towards the minimum displacement position, it is also possible that the radial sizes of the chambers progressively decrease, if the constructional or operating conditions demand a decrease in the torque exerted by vanes  423 ,  424  along the arc of rotation of ring  12 . In both cases, bases  425   a ,  426   a  might have non uniform curvatures (in any case, curvatures such that the radial size of the respective chamber is in the whole increasing or decreasing), so that a discontinuous variation of the active areas of vanes  423 ,  424 , and hence a discontinuously varying torque along the arc of rotation of ring  12 , may be obtained. Of course, at the discontinuity regions, the bases must be shaped so as to allow vane rotation in both directions. 
     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.