Abstract:
A variable capacity pump includes a control ring moveable within a pump chamber to alter the volumetric capacity of the pump. First and second control chambers individually receive pressurized fluid to create forces to bias the control ring in a predetermined direction. A return spring urges the control ring toward a maximum volumetric capacity pump position. The control ring connects and disconnects the second control chamber from a source of pressurized fluid based on a position of the control ring. Forces from the control chambers and the spring act in combination with one another or against one another and against the spring force to establish first and second equilibrium pressures based on a pressurized or vented condition of the second control chamber.

Description:
FIELD 
       [0001]    The present invention relates to variable displacement vane pumps. More specifically, the present invention relates to a variable displacement variable pressure vane pump system for mechanical systems such as internal combustion engines or automated transmissions. The present disclosure relates to an improved pump and control device for providing better control of the output of the variable capacity pump. More specifically, the present invention relates to a flow demand optimized control mechanism to control the output of a variable capacity pump at different operating conditions. 
       BACKGROUND 
       [0002]    Pumps for incompressible fluids, such as oil, are often variable capacity vane pumps. Such pumps include a moveable pump ring, which allows the rotor eccentricity of the pump to be altered to vary the capacity of the pump. 
         [0003]    Having the ability to alter the volumetric capacity of the pump to maintain an equilibrium pressure is important in environments such as automotive lubrication pumps, wherein the pump will be operated over a range of operating speeds. In such environments, to maintain a comparatively equilibrium pressure it is known to employ a direct, or indirect, feedback supply of the working fluid (e.g. lubricating oil) from the output of the pump to a control chamber adjacent the pump control ring, the pressure in the control chamber acting to move the control ring, against a biasing force, typically from a return spring, to alter the capacity of the pump. 
         [0004]    When the pressure at the output of the pump increases, such as when the operating speed of the pump increases, the increased pressure is applied to the control ring to overcome the bias of the return spring and to move the control ring to reduce the capacity of the pump, thus reducing the output volume and hence the pressure at the output of the pump, to continue to maintain a comparatively equilibrium pressure despite the change in operating conditions, (speed). 
         [0005]    Conversely, as the pressure at the output of the pump drops, such as when the operating speed of the pump decreases, the decreased pressure applied to the control chamber adjacent the control ring allows the biasing force, typically from a return spring, to move the control ring to increase the capacity of the pump, raising the output volume and hence pressure of the pump, to continue to maintain a comparatively equilibrium pressure despite the change in operating conditions. In this manner, a comparatively equilibrium pressure is obtained at the output of the pump over a range of operating conditions (speeds). 
         [0006]    The equilibrium pressure is determined by the area of the control ring against which the working fluid in the control chamber acts, the pressure of the working fluid supplied to the chamber and the bias force, typically generated by the return spring and the characteristics of the hydraulic system that the pump operates within. 
         [0007]    Conventionally, the equilibrium pressure is selected to be a pressure which is acceptable for the expected operating range of the engine and is thus somewhat of a compromise as, for example, the engine may be able to operate acceptably at lower operating speeds with a lower working fluid pressure than is required at higher engine operating speeds. In order to prevent undue wear or other damage to the engine, the engine designers will select an equilibrium pressure for the pump which meets the worst case (for example, high engine load or operating speed) conditions. Thus, at lower speeds, or lower engine loads, the pump will be operating at a higher capacity than necessary, wasting energy pumping the surplus, unnecessary, working fluid through the hydraulic system. 
         [0008]    It is desired to have a simple variable capacity vane pump that can provide at least two equilibrium pressures in reasonably compact pump housing. Some prior art solutions use a dual spring configuration, as shown for example in WO2013049929 A1. It may be desirable to achieve similar benefits by using simple hydraulic connections, without the need for additional components. 
       SUMMARY 
       [0009]    It is an object of the present invention to provide a novel variable displacement variable pressure vane pump which obviates or mitigates at least one disadvantage of the prior art. 
         [0010]    A variable capacity pump includes a control ring moveable within a pump chamber to alter the volumetric capacity of the pump. First and second control chambers individually receive pressurized fluid to create forces to bias the control ring in a predetermined direction. A return spring urges the control ring toward a maximum volumetric capacity pump position. The control ring connects and disconnects the second control chamber from a source of pressurized fluid based on a position of the control ring. Forces from the control chambers and the spring act in combination with one another or against one another and against the spring force to establish first and second equilibrium pressures based on a pressurized or vented condition of the second control chamber. 
         [0011]    In a first arrangement, the return spring acts against the combined force of the two control chambers to establish a lower equilibrium pressure. After the control ring has moved a predetermined amount, a simple feature in the control ring is configured to close the hydraulic passage that energizes the second control chamber and opens a passage to vent the second control chamber. The return spring then acts against the force of only the first control chamber, to establish a second, higher equilibrium pressure. 
         [0012]    In a second arrangement, the return spring acts against the force of a primary control chamber to establish a lower equilibrium pressure. After the control ring has moved a predetermined amount, a simple feature in the control ring is configured to open a hydraulic passage that energizes a second control chamber, acting against the force of the primary control chamber. The return spring and the force in the secondary control chamber then acts against the force in the first control chamber, and therefore establish a second, higher equilibrium pressure. 
         [0013]    In a third arrangement, similar to the first one presented, a third chamber is added on the control ring and connected to the supply of working fluid by an ON/OFF Solenoid Valve to produce two relatively parallel pressure curves. A high mode is provided when the third chamber is not pressurized and a low mode when the third chamber is pressurized. 
         [0014]    In a fourth arrangement, similar to the second one presented, a third chamber is added on the control ring and connected to the supply of working fluid by an ON/OFF Solenoid Valve to produce two relatively parallel pressure curves. A high mode is produced when the third chamber is not pressurized, and a low mode when the third chamber is pressurized. 
     
    
     
       DRAWINGS 
         [0015]    The drawings described herein are for illustrative purposes only of selected embodiments and not all possible implementations, and are not intended to limit the scope of the present disclosure. 
           [0016]      FIG. 1  is a partial plan view of a variable capacity pump constructed in accordance with the teachings of the present disclosure; 
           [0017]      FIGS. 2A-2D  show the pump at different eccentricity stages; 
           [0018]      FIG. 3  is a graph of the pressure output of the pump depicted in  FIGS. 2A-2D  versus the oil pressure demand of the mechanical system; 
           [0019]      FIG. 4  is a partial plan view of another variable capacity pump; 
           [0020]      FIGS. 5A-5D  show the pump of  FIG. 4  different eccentricity stages; 
           [0021]      FIG. 6  is a partial plan view of another variable capacity pump; 
           [0022]      FIGS. 7A-7D  show the pump of  FIG. 6  at different eccentricity stages; 
           [0023]      FIG. 8  is a graph of the pressure output of the pump shown in  FIGS. 7A-7D  versus the minimum and maximum oil pressure demand of a mechanical system; 
           [0024]      FIG. 9  is a partial plan view of another variable capacity pump; 
           [0025]      FIGS. 10A-10D  show the pump of  FIG. 9  at different eccentricity stages; and 
           [0026]      FIG. 11  is a partial plan view of a variable capacity pump including a pendulum slider mechanism. 
       
    
    
       [0027]    Corresponding reference numerals indicate corresponding parts throughout the several views of the drawings. 
       DESCRIPTION 
       [0028]    A variable capacity vane pump in accordance with an embodiment of the present invention is indicated generally at  20  in  FIG. 1 . Pump  20  includes a casing or housing  22  with a front face  24  which is sealed with a pump cover (not shown) and optionally a suitable gasket (not shown), to an engine (not shown) or the like, for which pump  20  is to supply pressurized working fluid. 
         [0029]    Pump  20  includes a drive shaft  28  which is driven by any suitable means, such as the engine or other mechanism to which the pump is to supply working fluid, to operate pump  20 . As drive shaft  28  is rotated, a pump rotor  32  located within a pump chamber  36  is driven by drive shaft  28 . A series of slidable pump vanes  40  rotate with rotor  32 , the outer end of each vane  40  engaging the inner circumferential surface of a pump control ring  44 , which forms the outer wall of pump chamber  36 . Pump chamber  36  is divided into a series of working fluid chambers  48 , defined by the inner surface of pump control ring  44 , pump rotor  32  and vanes  40 . 
         [0030]    Pump control ring  44  is mounted within housing  22  via a pivot pin  52  that allows the center of pump control ring  44  to be moved relative to the center of rotor  32 . As the center of pump control ring  44  is located eccentrically with respect to the center of pump rotor  32  and each of the interior of pump control ring  44  and pump rotor  32  are circular in shape, the volume of working fluid chambers  48  changes as the chambers  48  rotate around pump chamber  36 , with their volume becoming larger at the low pressure side (the left hand side of pump chamber  36  in  FIG. 1 ) of pump  20 , and smaller at the high pressure side (the right hand side of pump chamber  36  in  FIGS. 2A-2D ) of pump  20 . This change in volume of working fluid chambers  48  generates the pumping action of pump  20 , drawing working fluid from a pump inlet  50  and pressurizing and delivering it to a pump outlet  54 . 
         [0031]    By moving pump control ring  44  about pivot pin  52  the amount of eccentricity, relative to pump rotor  32 , can be changed to vary the amount by which the volume of working fluid chambers  48  change from the low pressure side of pump  20  to the high pressure side of pump  20 , thus changing the volumetric capacity of the pump. A return spring  56  engages a tab  55  of control ring  44  and housing  22  to bias pump control ring  44  to the position, shown in  FIG. 1 , wherein the pump has a maximum eccentricity. 
         [0032]    A first control chamber  61  is formed between pump housing  22 , pump control ring  44 , a seal  71  and a seal  72 , mounted on pump control ring  44  and abutting housing  22 . In the illustrated configuration, first control chamber  61  is in direct fluid communication with pump outlet  54  such that pressurized working fluid from pump  20  which is supplied to pump outlet  54  also fills first control chamber  61 . 
         [0033]    As will be apparent to those of skill in the art, first control chamber  61  need not be in direct fluid communication with pump outlet  54  and can instead be supplied from any suitable source of working fluid, directly or indirectly, such as from oil gallery in an automotive engine being supplied by pump  20 . 
         [0034]    A second control chamber  62  is formed between pump housing  22 , pump control ring  44 , seal  72  and a seal  73 , mounted on pump control ring  44  and abutting housing  22 . 
         [0035]    Second control chamber  62  is supplied with pressurized fluid via a feeding orifice  81  into the housing  22 , and located partially under the pump control ring  44 . Pressurized fluid for orifice  81  can be supplied either from pump outlet  54 , or other source of working fluid, such as an oil gallery in an automotive engine. A discharge passage  82  is located in the housing  22  and under the pump control ring  44  in communication with the pump inlet  50 . A channel or recess  83  extends across the width of control ring  44  in a direction perpendicular to a direction that the control ring moves. As shown in  FIGS. 2A-2D , feeding orifice  81 , discharge passage  82  and recess  83  are positioned and sized to create a pump pressure output versus speed as shown in  FIG. 3 . There are four distinctive steps, shown in  FIGS. 2A-2D , that generate the pump pressure output curve. 
         [0036]    In curve portion A-B 1 , both first control chamber  61  and second control chamber  62  are energized because the feeding orifice  81  is connected to second control chamber  62  and the discharge passage  82  is not connected, being completely covered by the pump control ring  44 . However, at low pump operating speeds, the force and consequently the turning moment around the pivot pin  52  created by the pressure build up in the two control chambers is insufficient to counter the force of the return spring  56 , and as such the pump remains at maximum eccentricity. 
         [0037]    In curve portion B 1 -C 1 , the pressure build up due to higher speeds of the pump has generated enough force, from the pressure in the two control chambers and consequently the turning moment, acting around the pivot pin  52  to exceed the force of the return spring  56 , which is providing an opposing turning moment acting around the pin to reduce the pump control ring eccentricity. In this phase, the slight movement of the control ring  44  has not yet opened the discharge passage  82  to second control chamber  62 , hence both control chambers are still working. 
         [0038]    Curve portion C 1 -D 1  represents a transition phase, where the movement of the pump control ring started in portion B 1 -C 1  has reached a point where the recess  83  is changing second control chamber  62  connections. Pressure feeding orifice  81  is closed and discharge passage  82  is opened, ultimately venting second control chamber  62 . As such, with a further increase in operating speed and pressures, only first control chamber  61  is energized and a new force balance is established around pivot pin  52 . The pressure from first control chamber  61  acts against the force generated by the return spring  56 . In this phase, the slight pressure increase in first control chamber  61  cannot move the control ring  44  and the pump eccentricity remains essentially constant. 
         [0039]    In curve portion D 1 -E 1 , the pressure within first control chamber  61  increases due to higher pump operating speeds to generate enough force from the pressure in the first control chamber  61 , acting as a turning moment, around the pivot pin  52  to exceed the force of the return spring  56 , which is providing an opposing turning moment around the pin. A reduction of the pump control eccentricity occurs. 
         [0040]    Another pump constructed according to the principles of the present disclosure is shown in  FIG. 4  and identified at reference number  20   a.  In this arrangement, two control chambers are located on opposite sides of the pivot pin  52   a,  and act against each other. The pump outlet  54   a  is connected to a pressure port  57   a  via a drilled internal channel within the housing  22   a.  In this arrangement, a first control chamber  61   a  is formed in the pump chamber  36   a , between pump control ring  44   a,  pump housing  22   a,  seal  71   a  and pivot pin  52   a , and when energized, it creates a force, acting as a turning moment around pivot pin  52   a,  opposite to the force of the return spring  56   a.  In the illustrated configuration, first control chamber  61   a  is supplied with pressurized fluid from engine oil gallery or pump outlet via a feeding channel  84   a.    
         [0041]    A second control chamber  62   a  is formed in the pump chamber  36   a,  between pump control ring  44   a,  pump housing  22   a,  seal  72   a  and pivot pin  52   a,  and when energized, it creates a force, acting as a turning moment, around pivot pin  52   a,  acting in the same direction as the force of the return spring  56   a.    
         [0042]    Second control chamber  62   a  is supplied with pressurized fluid via a feeding orifice  81   a  into the housing  22   a,  and located under the pump control ring  44   a.  Pressurized fluid for orifice  81   a  can be supplied either from pump outlet  54   a,  or other source of working fluid, directly or indirectly, such as an oil gallery in an automotive engine. A discharge passage  82   a  located in the housing  22   a  and partially under the pump control ring  44   a,  is in connection to the pump inlet  50   a.  A channel  83   a  is shaped as a blind recess having an opening at an edge of control ring  44   a  that extends along a surface of the control ring that slides relative to pump housing  22 . As shown in  FIGS. 5A-5D , pump  20   a  is equipped with feeding orifice  81   a,  discharge passage  82   a,  and connecting channel  83   a  in pump control ring  44   a  to create a pump pressure output as shown in  FIG. 3 . There are four distinctive steps, shown in  FIGS. 5A-5D , that generate that pump pressure output curve. 
         [0043]    In curve portion A-B 1 , first control chamber  61   a  is energized via feeding channel  84   a  and second control chamber  62   a  is not energized, since second control chamber  62   a  is vented to the inlet via discharge passage  82   a  and the connecting channel  83   a.  The feeding orifice  81   a  is not connected to second control chamber  62   a,  being completely covered by the pump control ring  44   a.  At low pump operating speeds, the force, acting as a turning moment, around the pivot pin  52   a  created by the pressure build up in first control chamber  61   a  is not sufficient to counter the force created by the return spring  56   a,  and as such the pump remains at maximum eccentricity. 
         [0044]    At curve portion B 1 -C 1 , the pressure build up due to higher operating speeds of the pump has generated enough force from first control chamber  61   a,  acting as a turning moment, around the pivot pin  52   a  to exceed the force of the return spring  56   a,  acting as an opposing turning moment, around the pin, determining a reduction of the pump eccentricity. In this phase, the slight movement of control ring  44   a  has not yet connected the feeding orifice  81   a  to the connecting channel  83   a,  hence only first control chamber  61   a  is still working. 
         [0045]    Curve portion C 1 -D 1  represents a transition phase, where the movement of the pump control ring started in portion B 1 -C 1  has reached a point where the control channel  83   a  is changing second control chamber  62   a  connections, by connecting pressure feeding orifice  81   a  with second control chamber  62   a  and closing the second control chamber  62   a  connection to discharge passage  82   a.  As such, with further increase in pump operating speed and pressures, both control chambers  61   a  and  62   a  are energized and a new force balance is established around pivot pin  52   a.  The pressure from first control chamber  61   a  acts against the force generated by the return spring  56   a  and second control chamber  62   a.    
         [0046]    At curve portion D 1 -E 1 , the pressure build up due to higher operating speeds of the pump has generated enough force from first control chamber  61   a,  acting as a turning moment, around the pivot pin  52   a  to exceed the force of the return spring  56   a  combined with the force from second control chamber  62   a,  determining a reduction of the pump eccentricity. 
         [0047]    It should be appreciated that the feeding orifice  81 , discharge passage  82 , and recess  83  described in relation to pump  20  and depicted in  FIG. 1  may alternatively be applied to pump  20   a  in lieu of feeding orifice  81   a , discharge passage  82   a  and recess  83   a.  It is also contemplated that the geometry incorporated to provide the passive control features of pump  20   a  may be applied to pump  20 . 
         [0048]    Another alternate variable capacity pump is presented in  FIG. 6  and identified as reference number  20   b.  Pump  20   b  is substantially similar to pump  20  shown in  FIG. 1 , to which a third control chamber  63   b  connected to an electrically controlled hydraulic solenoid valve  91   b  was added. Use of the third control chamber  63   b  provides the flexibility to generate either a high (A-B 1 -C 1 -D 1 -E 1 ) or a low (A-B 2 -C 2 -D 2 -E 2 ) pump pressure output in relation to operating speed as shown in  FIG. 8 . It may be beneficial to provide a pump operable to meet different demand requirements that may occur during the operation on an automobile engine. For example, many newer vehicles are selectively operable in a high load engine pressure demand mode, as well as the more traditional low load engine pressure demand mode. A pressure output may be required from the pump to provide lubricating and cooling oil to an auxiliary system such as an internal combustion engine piston cooling system. The high load engine pressure demand curve in  FIG. 8  may include a greater inflection in the pressure versus engine speed curve at a predetermined engine speed. One skilled in the art should appreciate that the present configuration of pump  20   b  equipped with third control chamber  63   b  and solenoid valve  91   b  provides a simple and cost effective solution to the requirement for substantially different pressure demand curves. In particular, it is contemplated that electrically controlled hydraulic solenoid valve  91   b  is an inexpensive on/off valve. It should also be appreciated that if greater control is required, the electrically controlled solenoid valve may be a proportional type operable to modulate the pressure in third control chamber  63   b  between the system pressure and either atmospheric pressure or pump inlet pressure. 
         [0049]    As presented in  FIG. 6 , first control chamber  61   b  is formed between pump housing  22   b,  pump control ring  44   b,  seal  71   b  and seal  72   b , mounted on pump control ring  44   b  and abutting housing  22   b.  In the illustrated configuration, first control chamber  61   b  is in direct fluid communication with pump outlet  54   b  such that pressurized working fluid from pump  20   b  which is supplied to pump outlet  54   b  also fills first control chamber  61   b.    
         [0050]    As will be apparent to those skilled in the art, first control chamber  61   b  need not be in direct fluid communication with pump outlet  54   b  and can instead be supplied from any suitable source of working fluid, directly or indirectly, such as from an oil gallery in an automotive engine being supplied by pump  20   b.    
         [0051]    Second control chamber  62   b  is formed between pump housing  22   b,  pump control ring  44   b,  seal  73   b  and seal  74   b,  mounted on pump control ring  44   b  and abutting housing  22   b.  Second control chamber  62   b  is supplied with pressurized fluid via a feeding orifice  81   b  into the housing  22   b,  and located partially under the pump control ring  44   b.  Pressurized fluid for orifice  81   b  can be supplied either from pump outlet  54   b,  or other source of working fluid, such as an oil gallery in an automotive engine. A discharge passage  82   b  located into the housing  22   b  and under the pump control ring  44   b,  is in connection to the pump inlet  50   b.    
         [0052]    Third control chamber  63   b  is formed between pump housing  22   b,  pump control ring  44   b,  seal  72   b  and seal  74   b  and is supplied in pressurized oil from the solenoid valve  91   b  via a feeding channel  85   b.  As shown in  FIGS. 7A-7D , pump  20   b  includes feeding orifice  81   b,  discharge passage  82   b  and recess  83   b  in the pump control ring  44   b,  designed and sized to create a pump pressure output as shown in  FIG. 8 . When third control chamber  63   b  is not energized with pressurized working fluid from the solenoid valve, the pump works in high mode, and generates the pressure curve A-B 1 -C 1 -D 1 -E 1  as shown in  FIG. 8 . There are four steps, shown in  FIGS. 7A-7D , that generate the high mode pump pressure output curve. 
         [0053]    In curve portion A-B 1 , both first control chamber  61   b  and second control chamber  62   b  are energized, because the feeding orifice  81   b  is connected to second control chamber  62   b  and the discharge passage  82   b  is not connected, being completely covered by the pump control ring  44   b.  At low pump operating speeds, the force, acting as a turning moment, around the pivot pin  52   b  created by the pressure build up in control chambers  61   b,    62   b  is not sufficient to counter the force created by the return spring  56   b,  which is acting around the pin as an opposing turning moment, and as such the pump remains at maximum eccentricity. 
         [0054]    In curve portion B 1 -C 1 , the counter pressure build up due to higher operating speeds of the pump has generated enough force from the two control chambers, acting as a turning moment, around the pivot pin  52   b  to exceed the force of the return spring  56   b,  acting as an opposing turning moment, around the pin to reduce of the pump eccentricity. In this phase, the slight movement of the control ring  44   b  has not yet opened the discharge passage  82   b  to second control chamber  62   b,  hence both control chambers are still working. 
         [0055]    Curve portion C 1 -D 1  represents a transition phase, where the movement of the pump control ring started in portion B 1 -C 1  has reached a point where the recess  83   b  is changing second control chamber  62   b  connections, by closing its pressure feeding orifice  81   b  and opening the discharge passage  82   b,  ultimately venting second control chamber  62   b.  As such, with a further increase in pump operating speed, system pressure and feeding pressures, only second control chamber  62   b  is energized and a new force balance is established around pivot pin  52   b,  the pressure from second control chamber  62   b  acting against the force generated by the return spring  56   b.    
         [0056]    At curve portion D 1 -E 1 , the pressure due to higher operating speeds of the pump has generated enough force from first control chamber  61   b , acting around the pivot pin  52   b  to exceed the force of the return spring  56   b  acting around the pin, causing a reduction of the pump eccentricity. 
         [0057]    Pressure curve A-B 2 -C 2 -D 2 -E 2  is generated in a similar fashion with the exception that solenoid valve  91   b  is energized to provide pressurized fluid to third control chamber  63   b  via feeding channel  85   b.  A force acting in an opposite direction to the spring force is applied when third control chamber  63   b  is pressurized. As such, the eccentricity of control ring  44   b  is reduced. An offset, low pressure output curve results. 
         [0058]    Another variable capacity pump  20   c  is depicted in  FIG. 9 . Pump  20   c  is substantially similar to pump  20   a  with the exception that a third control chamber  63   c  connected to an electrically controlled hydraulic solenoid valve  91   c  are included. Control of valve  91   c  allows pump  20   c  to generate either the high (A-B 1 -C 1 -D 1 -E 1 ) or low (A-B 2 -C 2 -D 2 -E 2 ) pump pressure output in relation to operating speed. As presented in  FIG. 9 , two control chambers are located on one side of the pivot pin  52   c,  while a third control chamber and the return spring  56   c  are on an opposite side of the pivot. The pump outlet  54   c  is connected to the pressure port  57   c  via a drilled internal channel within the housing  22   c.  Pump  20   c  includes first control chamber  61   c  formed in the pump chamber  36   c,  between pump control ring  44   c,  pump housing  22   c,  seal  71   c  and pivot pin  52   c,  and when energized, it creates a force, acting as a turning moment around pivot pin  52   c,  opposite to the force of the return spring  56   c.  In the illustrated configuration, first control chamber  61   c  is supplied with pressurized fluid from engine oil gallery or pump outlet via a feeding channel  84   c.    
         [0059]    A second control chamber  62   c  is formed in the pump chamber  36   c,  between pump control ring  44   c,  pump housing  22   c,  seal  72   c  and pivot pin  52   c,  and when energized, it creates a force, acting as a turning moment, around pivot pin  52   c,  acting in the same direction as the momentum created by the force of the return spring  56   c.    
         [0060]    Second control chamber  62   c  is supplied with pressurized fluid via a feeding orifice  81   c  into the housing  22   c,  and located under the pump control ring  44   c.  Pressurized fluid for orifice  81   c  can be supplied either from pump outlet  54   c,  or other source of working fluid, directly or indirectly, such as an oil gallery in an automotive engine. A discharge passage  82   c  located into the housing  22   c  and partially under the pump control ring  44   c,  is in connection to the pump inlet  50   c.    
         [0061]    A third control chamber  63   c  is formed between pump housing  22   c,  pump control ring  44   c,  seal  71   c  and seal  73   c  and is supplied in pressurized oil from the solenoid valve  91   c  via a feeding orifice  87   c.  As shown in  FIGS. 10A-10D , pump  20   c  includes feeding orifice  81   c,  discharge passage  82   c  and connecting channel  83   c  in the pump control ring  44   c.  Pump  20   c  is designed and sized to create a pump pressure output as shown in  FIG. 8 . When third control chamber  63   c  is not pressurized, pump  20   c  generates pump pressure output curve A-B 1 -C 1 -D 1 -E 1  as shown in  FIGS. 10A-10D . 
         [0062]    At curve portion A-B 1 , first control chamber  61   c  is energized and second control chamber  62   c  is not energized, since second control chamber  62   c  is vented to the inlet via discharge passage  82   c  and the connecting channel  83   c.  The feeding orifice  81   c  is not connected to second control chamber  62   c , being completely covered by the pump control ring  44   c.  At low pump operating speeds, the force, acting as a turning moment, around the pivot pin  52   c  created by the pressure build up in first control chamber  61   c  is not sufficient to counter the force created by the return spring  56   c,  and as such the pump remains at maximum eccentricity. 
         [0063]    At curve portion B 1 -C 1 , the pressure build up due to higher operating speeds of the pump has generated enough force from first control chamber  61   c,  acting as a turning moment, around the pivot pin  52   c  to exceed the force of the return spring  56   c,  acting as an opposing turning moment, around the pin, determining a reduction of the pump eccentricity. In this phase, the slight movement of the control ring  44   c  has not yet connected the feeding orifice  81   c  to the connecting channel  83   c,  hence only first control chamber  61   c  is still working. 
         [0064]    Curve portion C 1 -D 1  represents a transition phase, where the movement of the pump control ring started in portion B 1 -C 1  has reached a point where the control channel  83   c  is changing second control chamber  62   c  connections, by connecting pressure feeding orifice  81   c  with second control chamber  62   c  and closing the second control chamber  62   c  connection to discharge passage  82   c.  As such, with further increase in pump operating speed and pressures, both first and second control chambers  61   c,    62   c  are energized and a new force balance is established around pivot pin  52   c.  The pressure from first control chamber  61   c  acts against the force generated by the return spring  56   c  and the second control chamber  62   c.    
         [0065]    At curve portion D 1 -E 1 , the pressure build up due to higher operating speeds of the pump has generated enough force from the first control chamber  61   c,  acting as a turning moment, around the pivot pin  52   c  to exceed the force of the return spring  56   c  combined with the force from second control chamber  62   c,  determining a reduction of the pump eccentricity. 
         [0066]    Pressure curve A-B 2 -C 2 -D 2 -E 2  is generated in a similar fashion when solenoid valve  91   c  is emerged. Pressurized working fluid is provided to third control chamber  63   c  via the feeding orifice  87   c.    
         [0067]      FIG. 11  depicts another alternate pump identified at  20   d . Pump  20   d  is substantially similar to pump  20 , with the exception that the pumping members used to urge fluid from the inlet to the outlet are configured as a pendulum-slide cell instead of the vane arrangement previously described. Accordingly, like elements will retain their previously introduced reference numerals including a “d” suffix. Pump  20   d  includes an inner rotor  102  coupled to a plurality of pendulum slides  104  via an outer rotor  106 . Pendulum slides  104  are pivotally mounted to outer rotor  106 . Pendulum slides  104  are movable within radially extending slots  108  extending into inner rotor  102 . Inner rotor  102  together with pendulum slides  104  and outer rotor  106  define pumping chamber  110 . According to the rotational position of inner rotor  102 , outer rotor  106 , pumping chambers  110  serve as suction chambers or as pressure chambers for transferring fluid. It should be appreciated with either the outer rotor  106  or the inner rotor  102  may be a driven member of pump  20   d.    
         [0068]    The above-described configurations are intended to be examples and alterations and modifications may be effected thereto, by those of skill in the art, without departing from the scope of the present disclosure. 
         [0069]    Moreover, it will be obvious to those skilled in the art that additional control chambers can be configured on either side of the pivot pin and these could be passively controlled by additional similar features in the control ring and therefore responsive to movement of the control ring. One or more of the control chambers may be actively controlled by an electrically operated solenoid valve to optimize the volume and pressure output characteristics of a pump to suit a given application.