Patent Abstract:
A lubrication system for a power transmission device includes a variable displacement vane pump including a moveable control ring for varying the displacement of the pump. A linear actuator directly acts on the control ring for moving the control ring between maximum and minimum pump displacement positions. The linear actuator includes an electric motor for rotating a drive member. The drive member engages a driven actuator shaft to cause linear translation of the actuator shaft in response to rotation of the drive member. A control system includes a controller for signaling the actuator to extend or retract the actuator shaft to vary the pump displacement.

Full Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
     This application claims the benefit of U.S. Provisional Application No. 61/103,593, filed on Oct. 8, 2008. The entire disclosure of the above application is incorporated herein by reference. 
    
    
     FIELD 
     The present disclosure relates to variable displacement vane pumps. More specifically, the present invention relates to a variable displacement vane pump and system whose output flow is continuously variable and which can be selected independent of the operating speed of the pump. 
     BACKGROUND 
     Mechanical systems, such as internal combustion engines and automatic transmissions, typically include a lubrication pump to provide lubricating oil, under pressure, to many of the moving components and/or subsystems of the mechanical systems. In most cases, the lubrication pump is driven by a rotating component of the mechanical system and thus the operating speed and output of the pump varies with the operating speed of the mechanical system. The lubrication requirements of the mechanical system do not directly correspond to the operating speed of the mechanical system. 
     To deal with these differences, prior art fixed displacement lubricating pumps were generally designed to operate effectively at a target speed and a maximum operating lubricant temperature resulting in an oversupply of lubricating oil at most mechanical system operating. A pressure relief valve was provided to return the surplus lubricating oil back into the pump inlet or oil sump to avoid over pressure conditions in the mechanical system. In some operating conditions such as low oil temperatures, the overproduction of pressurized lubricating oil can be 500% of the mechanical system&#39;s needs. The result is a significant amount of energy being used to pressurize the lubricating oil which is subsequently exhausted through the relief valve. 
     More recently, variable displacement vane pumps have been employed as lubrication oil pumps. Such pumps generally include a control ring, or other mechanism, which can be operated to alter the volumetric displacement of the pump and thus its output at an operating speed. Typically, a feedback mechanism is supplied with pressurized lubricating oil from the output of the pump to alter the displacement of the pump to operate and to avoid over pressure situations in the engine throughout the expected range of operating conditions of the mechanical system. 
     While such variable displacement pumps provide some improvements in energy efficiency over fixed displacement pumps, they still result in a significant energy loss as their displacement is controlled, directly or indirectly, by the output pressure of the pump which changes with the operating speed of the mechanical system, rather than with the changing requirements of the lubrication system. Accordingly, such variable displacement pumps must still be designed to provide oil pressures which meet the highest expected mechanical system requirements, despite operating temperatures and other variables, even when the mechanical system operating conditions normally do not necessitate such high requirements. 
     Another variable displacement pump control system is described within U.S. Pat. No. 7,018,178. The control system includes an electrical solenoid coupled to a variable displacement pump for varying the displacement of the pump during engine operation. While an electric solenoid may provide an additional degree of pump control, several disadvantages from its use exist. In particular, a solenoid requires a continuous supply of current to keep it active through operation of the pump. The use of the electrical power offsets the benefit of controlling the pump to minimize the amount of time where the pump provides excess lubricant flow. Furthermore, the maximum force capability of the solenoid is limited by the size of the electromagnet and the current applied thereto. For certain applications, the size of the electromagnet required to provide the desired force may be prohibitive for packaging the solenoid within an automotive environment. Accordingly, a need exists for an improved lubrication system capable of producing a desired lubricant flow while minimizing the energy required to do so. 
     SUMMARY 
     This section provides a general summary of the disclosure, and is not a comprehensive disclosure of its full scope or all of its features. 
     A lubrication system for a power transmission device includes a variable displacement vane pump including a moveable control ring for varying the displacement of the pump. A linear actuator directly acts on the control ring for moving the control ring between maximum and minimum pump displacement positions. The linear actuator includes an electric motor for rotating a drive member. The drive member engages a driven actuator shaft to cause linear translation of the actuator shaft in response to rotation of the drive member. A control system includes a controller for signaling the actuator to extend or retract the actuator shaft to vary the pump displacement. 
     Furthermore, a lubrication system for a power transmission device includes a variable displacement vane pump having a pivotable pump control ring for varying the displacement of the pump. A control system is operable to vary the displacement of the pump during operation of the pump to achieve an output pressure selected from a continuously variable range of output pressures from the pump which are independent from the operating speed of the pump. The control system includes a linear actuator coupled to the control ring for moving the control ring between minimum and maximum pump displacement positions. The linear actuator includes an electric stepper motor for bi-directionally rotating a nut threadingly engaged with an axially moveable actuator shaft. A coupler interconnects the shaft and the control ring and has multiple degrees of freedom to allow concurrent axial movement of the actuator shaft and rotation of the control ring. 
     Further areas of applicability will become apparent from the description provided herein. The description and specific examples in this summary are intended for purposes of illustration only and are not intended to limit the scope of the present disclosure. 
    
    
     
       DRAWINGS 
       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. 
         FIG. 1  is a cross-sectional view of an exemplary directly controlled variable displacement vane pump; 
         FIG. 2  is a sectional view of a portion of the pump and actuator assembly shown in  FIG. 1 ; 
         FIG. 3  is an enlarged fragmentary perspective view of the pumping system depicted in  FIGS. 1 and 2 ; 
         FIG. 4  is a schematic of an open loop control system for controlling the variable displacement vane pump; 
         FIG. 5  is a schematic depicting a closed loop control system cooperating with the variable displacement vane pump; 
         FIG. 6  is a fragmentary perspective view of an alternate connector coupling the actuator shaft and the control ring; 
         FIG. 7  is a fragmentary perspective view of another alternate connector coupling the actuator shaft and the control ring; 
         FIG. 8  is a fragmentary perspective view of another alternate connector coupling the actuator shaft and the control ring; 
         FIG. 9  is a fragmentary perspective view of another alternate connector coupling the actuator shaft and the control ring; 
         FIG. 10  is a fragmentary perspective view of another alternate connector coupling the actuator shaft and the control ring; 
         FIG. 11  is a fragmentary perspective view of another alternate connector coupling the actuator shaft and the control ring; 
         FIG. 12  is a fragmentary perspective view of another alternate connector coupling the actuator shaft and the control ring; 
         FIG. 13  is a sectional view of another alternate connector coupling the actuator shaft and the control ring; and 
         FIG. 14  is a fragmentary perspective view of another alternate connector coupling the actuator shaft and the control ring. 
     
    
    
     Corresponding reference numerals indicate corresponding parts throughout the several views of the drawings. 
     DETAILED DESCRIPTION 
     Example embodiments will now be described more fully with reference to the accompanying drawings. 
     With reference to  FIGS. 1-3 , a pumping system  10  is shown plumbed in communication with an exemplary power transmission device  12 . Power transmission device  12  is shown schematically and may include any number of devices including an internal combustion engine, a transmission, a transfer case, an axle assembly or the like. Pumping system  10  includes a variable displacement pump  14  including a housing  16  with a flange  17  for mounting pump  14  to power transmission device  12 . Alternatively, housing  16  may be integrally formed with the power transmission device. An inlet  18  extends through housing  16  interconnecting a low pressure gallery  20  with a sump  22  storing the fluid to be pumped. An outlet  24  of housing  16  interconnects a high pressure chamber  26  with power transmission device  12 . 
     Pump  14  includes a pump rotor  28  rotatably mounted within a rotor chamber  32 . A drive shaft  34  is fixed for rotation with pump rotor  28  to provide energy for pumping the lubricant. A plurality of pump vanes  36  are coupled to rotor  28  and radially slidable relative thereto. The radial outer end of each vane  36  engages an inner surface  38  of a pump control ring  40 . A plurality of pumping chambers  44  are defined by inner surface  38 , pump rotor  28  and vane  36 . Control ring  40  includes an integrally formed pivot pin  46  positioned within a recess  48  formed in housing  16 . It should be appreciated that control ring  40  may be pivotally mounted within housing  16  via many other suitable methods as well. Inner surface  38  of pump control ring  40  has a circular cross-sectional shape. An outer surface  50  of rotor  28  also has a circular cross-sectional shape. The center of surface  38  is eccentrically located with respect to the center of surface  50 . Accordingly, the volume of each pumping chamber  44  changes as rotor  28  rotates. The volume of chambers  44  increases at the low pressure side of the pump in communication with inlet  18 . Pumping chambers  44  decrease in size at the high pressure side in communication with outlet  24  of pump  14 . The change in volume of pumping chambers  44  generates the pumping action by drawing working fluid from sump  22  and delivering pressurized fluid from outlet port  24 . 
     The output of pump  14  may be varied by rotating pump control ring  40  about pivot pin  46 . In particular, the amount of eccentricity between inner surface  38  of pump ring  40  and the outer surface  50  of rotor  28  changes as control ring  40  is rotated. 
     A radially outwardly protruding arm  60  is integrally formed with control ring  40  and protrudes outside of pumping chambers  44 . An actuator assembly  62  is coupled to arm  60  and is operable to move control ring  40  between a first position, a second position and any point therebetween. In the first position, the control ring provides maximum eccentricity and maximum pump flow. At the second position, control ring  40  is positioned at a minimum eccentricity relative to rotor  28  and a minimum of output occurs. 
     To reduce the magnitude of force required to be provided by actuator assembly  62 , a first pressure balance chamber  64  is formed on a first side of control ring  40  while a second pressure balance chamber  66  is formed on an opposite side of control ring  40 . First pressure balance chamber  64  and second pressure balance chamber  66  are each in fluid communication with pressurized fluid provided from outlet  24 . This arrangement effectively balances the forces acting on control ring  40  thereby minimizing the force required to move control ring  40  and vary the pump output. It should be appreciated that the pressure balanced arrangement may be desirable but is not a requisite portion of pumping system  10 . With the pressure balancing chambers, actuator  62  may function but may be tasked to provide a greater input force to move control ring  40 . 
     Actuator assembly  62  includes an electric stepper motor  70  including a stator  72  and a rotor  74  supported in a housing  75 . Rotor  74  is coupled to a nut  76  that is threadingly engaged with an externally threaded actuator shaft  78 . Housing  75  includes a flange  79  coupled to pump housing  16 . Flange  79  may alternatively be fixed to power transmission device  12 . Actuator shaft  78  includes a distal end  80  coupled to arm  60  by a connector  81 . A yoke  82  includes a first end  84  rotatably coupled to arm  60  via a pin  86 . A second end  88  of yoke  82  is bifurcated defining a slot  90  bounded by first and second fingers  92 ,  94 . A clevis pin  96  rotatably interconnects yoke  82  and actuator shaft  78 . 
     Referring to  FIG. 4 , actuator assembly  62  is in communication with a controller  100 , a power supply  102  and a drive  104 . Controller  100  may be programmed with an algorithm or algorithms referencing speed, pressure, flow or temperature maps to enable the controller to control the flow of the pump using an open loop control system as depicted in  FIG. 4 .  FIG. 5  depicts a closed loop control system including a pressure sensor  106  in communication with controller  100 . 
     In operation, driveshaft  34  begins to rotate and drive rotor  28 . Lubricant pressure and flow begin to increase at outlet  24 . At start-up, controller  100  locates control ring  40  in the first position. As such, flow increases linearly with the speed of driveshaft  34 . At a particular speed, the flow produced by pump  14  will exceed the lubrication requirements of power transmission device  12 . At this time, controller  100  provides a signal to drive  104 . Drive  104  is in receipt of electrical power from power supply  102 . Drive  104  generates electrical pulses and supplies pulses to electric stepper motor  70  causing nut  76  to rotate in one of two directions to extend or retract actuator shaft  78  as signaled by controller  100 . Because actuator shaft  78  is directly coupled to control ring  40 , the linear motion of actuator shaft  78  changes the eccentricity of the pump and thus the pump output flow. 
     When the open loop control system of  FIG. 4  is implemented, controller  100  continues to signal drive  104  to position control ring  40  based on any one or more of speed, pressure, flow or temperature mappings of the control algorithm. A dedicated pressure sensor associated with pump  14  is not required. Alternatively, the closed loop feedback system depicted in  FIG. 5  includes pressure sensor  106  providing a signal indicative of the pressure output by pump  14  to controller  100 . Controller  100  outputs a signal to drive  104  to position control ring  40  and cause pump  14  to output a desired lubricant pressure. 
       FIG. 6  depicts an alternate method of drivingly interconnecting actuator shaft  78  and arm  60 . A threaded sleeve  110  includes a threaded throughbore  112 . Actuator shaft  78  is threadingly engaged with threaded bore  112 . A connector  114  includes a first end having a reduced diameter and an externally threaded portion  116  as well as another portion  118  including a transversely extending through aperture. Threaded portion  116  is engaged with threaded bore  112  to fix threaded sleeve  110  to connector  114 . An elongated slot  120  extends through arm  60  in a direction substantially perpendicular to the direction of travel of actuator shaft  78 . A pin  122  extends through slot  120  and the aperture formed in connector  114  to drivingly interconnect actuator shaft  78  and control ring  40  while allowing the requisite degrees of freedom to allow control ring  40  to rotate while actuator shaft  78  linearly translates. 
       FIG. 7  depicts another alternate method of interconnecting actuator shaft  78  and control ring  40 . A driver  130  includes one end having an internally threaded bore  132  and an opposite end having a substantially spherical outer surface  134 . Threaded bore  132  is coupled to an externally threaded end  136  of actuator shaft  78 . Arm  60  includes a cam surface  138  engaged by spherical surface  134  of driver  130 . A spring  140  is positioned within a cavity  142  shown in  FIG. 1 . Spring  140  biases arm  60  into engagement with spherical surface  134 . In this manner, a constant engagement between surface  138  and spherical surface  134  will be maintained throughout operation of pumping system  10 . Furthermore, spring  140  urges control  40  toward the position of maximum eccentricity. 
     With reference to  FIG. 8 , another alternate method for interconnecting actuator shaft  78  and control ring  40  is illustrated. A clevis  150  includes a threaded internal bore  152  fixed to an externally threaded portion of actuator shaft  78 . Clevis  150  includes a bifurcated end opposite threaded bore  152  including a first leg  154  spaced apart from a second leg  156 . A connector  158  includes a first end  160  positioned between first leg  154  and second leg  156 . A first arm  164  and a second arm  166  are integrally formed with control ring  40 . A second end  162  of connector  158  is positioned between first and second arms  164 ,  166 . A pin  168  interconnects connector  158  with control ring  40  and allows relative rotation therebetween. Once clevis  150  is threadingly engaged with actuator shaft  78  and connector  158  is pinned to control ring  40 , connector  158  is rotated in alignment with clevis  150  to allow insertion of another pin  170  rotatably interconnecting connector  158  to clevis  150 . 
     Another alternate interconnection method is shown in  FIG. 9 . A clevis  180  includes an open frame portion  182  having a through aperture  184  extending through one portion of the frame. An opposite portion of the frame includes integrally formed and spaced apart first and second legs  186 ,  188 . A distal portion of actuator shaft  78  extends through aperture  184 . A nut  190  threadingly engages an externally threaded portion of actuator shaft  78  to fix clevis  180  to actuator shaft  78 . A connector  192  includes a cylindrically shaped portion  194  and a radially protruding shaft portion  196 . A flattened portion  198  is formed at the distal end of shaft portion  196  and positioned between first and second legs  186 ,  188 . A pin  200  rotatably interconnects connector  192  and clevis  180 . Cylindrical portion  194  is rotatably coupled to control ring  40  by being positioned within a cylindrically shaped seat  202  of an integrally formed arm  204 . Shaft portion  196  extends through a slot  206  formed in arm  204 . 
       FIG. 10  depicts another method of interconnecting actuator shaft  78  and control ring  40 . In particular, a ball joint assembly  210  and a connector  212  couple actuator shaft  78  to a bifurcated pair of arm portions  214 ,  215  integrally formed with control ring  40 . Ball joint assembly  210  includes a socket  216  having a first end fixed to actuator shaft  78  and a second end defining a substantially spherical concave surface  220 . Ball joint assembly  210  also includes a ball stud  222  including a shank  224  and a ball  226  integrally formed with each other. Ball  226  engages spherical surface  220  of socket  216 . Connector  212  is threadingly engaged with shank  224  and positioned between arms  214 ,  215 . A pin  228  rotatably interconnects connector  212  and control ring  40 . 
       FIG. 11  depicts a similar connection system to that described in relation to  FIG. 10 . Accordingly, like elements will retain their previously introduced reference numerals including an “A” suffix. The connection system of  FIG. 11  eliminates connector  212 A and utilizes pin  228 A to rotatably interconnect shank  224 A and control ring  40 . 
       FIG. 12  shows another connection including a ball joint assembly  230  including a socket  232  fixed to actuator shaft  78  and a ball shank  234  fixed to a clevis  236 . Ball shank  234  may be coupled to clevis  236  via a threaded interconnection or another load transferring method. Clevis  236  includes a bifurcated end  237  coupled for rotation with arm  60  by a pin  238 . 
     As shown in  FIG. 13 , another method of drivingly interconnecting actuator shaft  78  and a control ring  239  is depicted. In this arrangement, a ball stud  240  is fixed to the distal end of actuator shaft  78 . Control ring  239  includes an integrally formed pocket having a cylindrically shaped surface  244 . The cylindrical surface  244  extends an arc length greater than 180 degrees to retain a spherically shaped ball  246  of ball stud  240  therein. Surface  244  extends substantially the entire width of control ring  239  to allow ball stud  240  to be inserted within the recess prior to interconnection to actuator shaft  78 . Conversely, ball stud  240  may be fixed to actuator shaft  78  and then subsequently coupled to control ring  239 . 
     Yet another method for interconnecting actuator shaft  78  and control ring  40  is depicted at  FIG. 14 . A ball joint assembly  250  and an adapter  252  couple actuator shaft  78  to control ring  40 . One end of adapter  252  is fixed to a distal end of actuator shaft  78  via a threaded connection. An opposite end of adapter  252  is coupled to a socket  254  of ball joint assembly  250  via another threaded interconnection. A ball stud  256  extends between bifurcated arms  258 ,  260  of control ring  40 . A pin  262  rotatably interconnects ball shank  256  with control ring  40 . 
     A number of coupling techniques have been described to facilitate a ridged mounting of actuator housing  75  to pump housing  16  or another portion of power transmission device  12 . The connection provides sufficient degrees of freedom to allow actuator shaft  78  to linearly translate and transfer a force to the pivotally moveable control ring  40 . While many of the interconnections have been described as threaded couplings, it should be appreciated that any number of methods for fixing two components relative to one another such as pinning, riveting, welding, press-fitting, adhesive bonding or the like, are contemplated as being within the scope of the present disclosure. Furthermore, while the closed loop control system was previously described as being in communication with a pressure sensor, it should be appreciated that any number of other sensors may be implemented to provide controller  100  with data for decision making relating to the control of actuator  62  and pumping system  10 . 
     Furthermore, the foregoing discussion discloses and describes merely exemplary embodiments of the present disclosure. One skilled in the art will readily recognize from such discussion, and from the accompanying drawings and claims, that various changes, modifications and variations may be made therein without departing from the spirit and scope of the disclosure as defined in the following claims.

Technology Classification (CPC): 5