Patent Publication Number: US-6216670-B1

Title: Hydraulically-actuated system having a variable delivery fixed displacement pump

Description:
RELATION TO OTHER PATENT APPLICATION 
     This application is a continuation-in-part of application Ser. No. 09/038,121, filed Mar. 11, 1998 and entitled Hydraulically-Actuated System Having a Variable Delivery Fixed Displacement Pump, now U.S. Pat. No. 6,035,828. 
    
    
     TECHNICAL FIELD 
     The present invention relates generally to pumps and hydraulically-actuated systems used with internal combustion engines, and more particularly to a variable delivery fixed displacement pump and its associated system. 
     BACKGROUND ART 
     U.S. Pat. No. 5,515,829 to Wear et al. describes a variable displacement actuating fluid pump for a hydraulically-actuated fuel injection system. In this system, a high pressure common rail supplies pressurized lubricating oil to a plurality of hydraulically-actuated fuel injectors mounted in a diesel engine. The common rail is pressurized by a variable displacement swash plate type pump that is driven directly by the engine. Pressure in the common rail is controlled in a two-fold manner. First, some pressure control is provided by electronically varying the swash plate angle within the pump. However, because variable angle swash plate type pumps typically have a relatively narrow band of displacement control, pressure in the common rail is primarily controlled through an electronically controlled pressure regulator. The pressure regulator returns a portion of the pressurized fluid in the common rail back to the low pressure fluid sump in order to maintain fluid pressure in the common rail at a desired magnitude. 
     While the Wear et al. hydraulically-actuated system using a variable displacement pump has performed magnificently for many years in a variety of diesel engines manufactured by Caterpillar, Inc. of Peoria, Ill., there remains room for improvement. On the overall level, the Wear et al. system is relatively more complex in that the control scheme in its electronic control module must simultaneously control both the angle of the swash plate within the high pressure pump and the amount of fluid spilled via the pressure regulator. Also, variable angle swash plate type pumps are relatively complex, and thus more prone to mechanical breakdown relative to simple fixed displacement type pumps. Finally, the Wear et al. system inherently wastes energy that inevitably results in a higher than necessary fuel consumption for the engine. In other words, energy is wasted each time the pressure regulator spills an amount of pressurized fluid back to the low pressure sump. 
     In another class of fuel injection systems, an individual unit pump supplies pressurized fuel to each individual fuel injector. In still another class of fuel injection systems, an individual unit pump supplies actuating fluid to each individual hydraulically-actuated fuel injector. In both of these classes of fuel injection systems, the individual unit pumps are of a fixed displacement and fixed output type such that a substantial amount of energy is wasted in those instances where an amount of fuel injected corresponds to less than the full stroke of the individual unit pumps. In most instances, these fuel injection systems utilize electronically controlled spill valves on the fuel injectors themselves in order to control injection timing and/or quantity. However, because of space constraints at and around the engine head, compromises must often be made in order to accommodate such a spill valve in or on the fuel injector body, as well as the associated plumbing, etc. 
     The present invention is directed to overcoming these and other problems associated with, and improving upon, fluid pumps and their associated hydraulically-actuated systems. 
     DISCLOSURE OF THE INVENTION 
     A hydraulically-actuated system includes a fixed displacement pump with at least one piston that reciprocates in a pump housing that defines a high pressure area and a low pressure area. A control valve is attached to the pump housing and is moveable between a first position in which the piston displaces fluid in a first proportion between the high pressure area and low pressure area, and a second position in which piston displaces fluid in a second proportion between the high pressure area and the low pressure area. At least one hydraulically-actuated device is fluidly connected to the high pressure area, and a source of low pressure fluid is connected to the low pressure area of the pump. An electronic control module is in communication with and capable of controlling a position of the control valve. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     FIG. 1 is a schematic illustration of a hydraulically-actuated system according to the present invention. 
     FIG. 2 is a sectioned side diagrammatic view of a fixed displacement pump according to one aspect of the present invention. 
     FIG. 3 is a schematic illustration of the fluid plumbing for one piston of the fixed displacement pump of FIG.  2 . 
     FIGS. 4 a  and  4   b  are schematic illustrations of the sleeve metering control feature for the fixed displacement pump of FIG.  2 . 
     FIG. 5 is an enlarged side sectioned diagrammatic view of a control valve for controlling the delivery output of the fixed displacement pump of FIG.  2 . 
     FIGS. 6 a-d  are graphs of solenoid current fluid pressure, poppet valve position and sleeve position, respectively, versus time for the hydraulically-actuated system of the present invention. 
     FIG. 7 is a schematic illustration of fixed displacement variable output unit pump according to the present invention. 
    
    
     BEST MODE FOR CARRYING OUT THE INVENTION 
     Referring now to FIG. 1, a hydraulically actuated system  10  is attached to an internal combustion engine  9 . The hydraulic system includes a high pressure common fluid rail  12  that supplies high pressure actuation fluid to a plurality of hydraulically-actuated devices, such as hydraulically-actuated fuel injectors  13 . Those skilled in the art will appreciate that other hydraulically-actuated devices, such as actuators for gas exchange valves for exhaust brakes, could be substituted for the fuel injectors  13  illustrated in the example embodiment. Common rail  12  is pressurized by a variable delivery fixed displacement pump  16  via a high pressure supply conduit  19 . Pump  16  draws actuation fluid along a low pressure supply conduit  20  from a source of low pressure fluid  14 , which is preferably the engine&#39;s lubricating oil sump. Although other available liquids could be used, the present invention preferably utilizes engine lubricating oil as its hydraulic medium. After the high pressure fluid does work in the individual fuel injectors  13 , the actuating fluid is returned to sump  14  via a drain passage  25 . 
     As is well known in the art, the desired pressure in common rail  12  is generally a function of the engine&#39;s operating condition. For instance, at high speeds and loads, the rail pressure is generally desired to be significantly higher than the desired rail pressure when the engine is operating at an idle condition. An operating condition sensor  23  is attached to engine  9  and periodically provides an electronic control module  15  with sensor data, which includes engine speed and load conditions, via a communication line  24 . In addition, a pressure sensor  21  periodically provides electronic control module  15  with the measured fluid pressure in common rail  12  via a communication line  22 . The electronic control module  15  compares a desired rail pressure, which is a function of the engine operating condition, with the actual rail pressure provided by pressure sensor  21 . 
     If the desired and measured rail pressures are different, the electronic control module  15  commands movement of a control valve  17  via a communication line  18 . The position of control valve  17  determines the amount of fluid that leaves pump  16  via high pressure supply conduit  19  to high pressure rail  12 . Both control valve  17  and pump  16  are preferably contained in a single pump housing  30 . Unlike prior art hydraulic systems, the present invention controls pressure in common rail  12  by controlling the delivery output from pump  16 , rather than by wasting energy through the drainage of pressurized fluid from common rail  12  in order to achieve a desired pressure. 
     Referring now to FIGS. 2-4, the various features of pump  16  are contained within a pump housing  30 . Pump  16  includes a rotating shaft  31  that is coupled directly to the output of the engine, such that the rotation rate of shaft  31  is directly proportional to the drive shaft of the engine. A fixed angle swash plate  33  is attached to shaft  31 . The rotation of swash plate  33  causes a plurality of parallel disposed pistons  32  to reciprocate from left to right. In this example, pump  16  includes five pistons  32  that are continuously urged toward swash plate  33  by individual return springs  46 . Return springs  46  maintain shoes  34 , which are attached to one end of each piston  32  in contact with swash plate  33  in a conventional manner. Because swash plate  33  has a fixed angle, pistons  32  reciprocate through a fixed reciprocation distance with each rotation of shaft  31 . Thus, pump  16  can be thought of as a fixed displacement pump; however, control valve  17  determines whether the fluid displaced is pushed into a high pressure area past check valve  37  or spilled back into a low pressure area  36  via a spill port  35 . 
     The proportion of fluid displaced by pistons  32  to the respective high pressure area  40  (see FIG. 3) and low pressure area  36  within pump housing  30  is determined by the position of individual sleeves  51  that are mounted to move on the outer surface of the individual pistons  32 . Each sleeve  51  is connected to move with a central actuator shaft  50  via an annulus  52 . An actuator biasing spring  61  normally biases actuator shaft  50  toward the left to a position in which virtually all the fluid displaced by the individual pistons  32  escapes back into low pressure area  36  via spill port  35 . 
     Pressure within pumping chamber  39 , under each piston  32 , can only build when internal passage  42  and spill port  35  are covered by a sleeve  51 . When sleeve  51  covers spill port  35 , fluid displaced by piston  30  is pushed past check valve  37 , into a high pressure connecting annulus  40  and eventually out of outlet  41  to the high pressure rail  12 . When pistons  32  are undergoing the retracting portion of their stroke due to the action of return spring  46 , low pressure fluid is drawn into pumping chamber  39  from a low pressure area  36  within pump housing  30  past inlet check valve  38 . 
     Referring now specifically to FIGS. 4 a  and  4   b,  the internal passage  42  within each piston  32  extends between its pressure face end  43  and its side surface  44 . In this embodiment, the height of the individual sleeves  51  is about equal to the fixed reciprocation distance  45  of pistons  32 . In this way, when sleeve  51  is in the position shown in FIG. 4 a,  all of the fluid displaced by piston  32  is pushed into the high pressure area  40  (FIG. 3) within the pump  16 . On the other hand, when sleeve  51  is in the position shown in FIG. 4 b,  virtually all of the fluid displaced by piston  32  is spilled back into low pressure area  36  (FIGS. 2 and 3) within pump  16  via internal passage  42  and spill port  35 . Thus, pump  16  can be characterized as variable delivery since the high pressure output is variable, but also be characterized as a fixed displacement swash plate type pump since the pistons always reciprocate a fixed distance. 
     Referring now to FIG. 5, the internal structure of control valve  17 , which controls the position of sleeves  51 , is illustrated. Control valve  17  includes a linear actuator  70  that includes a solenoid armature  71 , a stator  72 , and a solenoid coil  74 . A poppet valve member  73  is moved downward toward valve seat  62  when current is supplied to solenoid coil  74 . Thus, when current is high, poppet valve member  73  is seated in valve seat  62  to close fluid communication between control volume  60  and a low pressure area  63 , which is in fluid communication with a low pressure passage  64 . When current is lower, fluid pressure in control volume  60  pushes poppet valve member  73  and armature  71  upward to open some fluid communication between control volume  60  and low pressure area  63  past valve seat  62 . 
     As stated earlier, actuator shaft  50  is normally biased downward by a biasing spring  61 . In addition to this spring force, actuator shaft  50  has a pair of opposing hydraulic surfaces that provide the means by which actuator shaft  50 , and hence sleeves  51  are moved and stopped between the respective positions shown in FIGS. 4 a  and  4   b.  In particular, actuator shaft  50  includes a shoulder area  53  that is always in fluid communication with the high pressure area within pump  16  via a high pressure conduit  54 . This high fluid pressure in conduit  54  is channeled via a central restricted communication passage  55  into control volume  60 . Fluid pressure in control volume  60  acts on a control pressure surface  56 , which is preferably about equal to the hydraulic surface area defined by shoulder area  53 . Thus, when fluid pressure in control volume  60  is equal to the high pressure in conduit  54 , the only force acting on actuator shaft  50  comes from biasing spring  61 . This occurs when current to solenoid coil  70  is high such that poppet valve member  73  is pushed to close fluid flow past valve seat  62 . When current to solenoid coil  74  is turned off, poppet valve member  73  is pushed off of valve seat  62  and the resulting fluid flow into low pressure area  63  lowers pressure in control volume  60  sufficiently that actuator shaft  50  has a tendency to move completely upward under the action of the high fluid pressure force acting on shoulder area  53 . The pressure in control volume  60 , and hence the position of actuator shaft  50  can be controlled to stop at any position depending upon the magnitude of the current being supplied to solenoid current  74 . Thus, depending upon the current to solenoid coil  74 , the amount of fluid pumped into the high pressure rail can be varied from zero to the maximum output of the pump. In the event of an electrical malfunction, over-pressurization of the rail is prevented since the actuator shaft  50  is biased downward by spring  61  where no high pressure output is produced. 
     Referring now to FIG. 7, a unit pump  116  version of the present invention is illustrated. In this embodiment, a cam  112  rotates to drive the reciprocation of a piston  132  that is at least partially positioned within a pump housing  130 . The pump housing  130  defines a low pressure area  136  that includes an inlet  147  connected to a source of low pressure fluid  114  via a low pressure supply line  120 . The pump housing  130  also defines a high pressure area  140  that includes an outlet  141  fluidly connected to a hydraulically-actuated device  113  via a high pressure supply line  119 . The piston  132  and the pump housing  130  define a pump chamber  139  that is fluidly connected to the low pressure area  136  and the high pressure area  140  past respective check valves  138  and  139  in a conventional manner. Piston  132  is biased toward a retracted position to follow the contour of cam  112  by a return spring  146 . As with the previous embodiment, piston  130  reciprocates through a fixed distance and thus displaces a fixed amount of fluid with each reciprocation. However, the relative proportions of the fluid displaced to high pressure area  140  and low pressure area  136  is controlled by the positioning of a sleeve  151 . When sleeve  151  is in the position shown, virtually all of the fluid displaced by the movement of piston  132  is displaced into low pressure area  136  due to the fluid connection between pumping chamber  139  via internal passage  142  and spill port  135 . The positioning of sleeve  151  is controlled via a suitable mechanical and/or hydraulic linkage to a control valve  117 , which can be of a type described earlier. In other words, control valve  117  is controlled in its position via an electronic control module  115  via a communication line  122  in a conventional manner. 
     INDUSTRIAL APPLICABILITY 
     Referring now in addition to FIGS. 6 a-d,  the operation of hydraulically-actuated system  10  will be described and illustrated. FIGS. 6 a  and  6   b  illustrate that the steady state rail pressure is directly proportional to the steady state current being supplied to the solenoid portion of control valve  17 . When solenoid current is low, rail pressure remains low. When solenoid current is high, rail pressure is raised accordingly. A medium current puts the rail pressure at a medium magnitude. The variation in solenoid current changes the amount of fluid being spilled past valve seat  62  which changes the fluid pressure in control volume  60 . With each change in fluid pressure within control volume  60 , actuator shaft  50  will seek out a new equilibrium position in which the hydraulic force acting on shoulder area  53  is balanced against the combined force from spring  61  and the hydraulic force acting on control pressure surface  56 . 
     Of interest in FIGS. 6 a - 6   d  is when the system is commanded to raise rail pressure. When this occurs, solenoid current jumps and the poppet valve member is driven to close valve seat  62 . This in turn causes actuator shaft  50  to move all the way to the left such that the complete stroke of the piston is utilized to pressurize fluid. This causes a rapid rise in rail pressure. When it is desired to lower the rail pressure, current to the solenoid is decreased. This quickly causes actuator shaft  50  to move to the right where the pistons have no effective pumping force. Pressure in the rail quickly drops as the hydraulically-actuated devices  13  continue to operate and consume the pressurized fluid in the common rail  12 . 
     The present invention decreases the complexity of prior art hydraulically-actuated systems by having only one electronically-controlled device for controlling pressure in the high pressure rail. Recalling in the prior art, two different control schemes were necessary as one controlled the swash plate angle in the pump and the other controlled the pressure regulator attached to the high pressure rail. The present invention accomplishes the same task by only controlling high pressure output from the pump. The present invention also improves the robustness of the hydraulically-actuated system since fixed angle swash plate type pumps are generally more reliable and less complex than the variable angle swash plate type pumps of the prior art. In addition, only one electronically-controlled actuator is utilized in the present invention. Finally, the overall fuel consumption of the engine utilizing the present invention should be improved over that of the prior art since the pump only pressurizes an amount of fluid that is actually used by the hydraulic devices, and therefore almost no energy is wasted. Recalling that in the case of the prior art, pressure in the common rail was maintained at least in part by returning an amount of pressurized fluid back to the sump, which resulted in an efficiency drop and waste of energy. 
     Referring again to FIG. 7, when in operation in a hydraulic system, the unit pump  116  has the ability to deliver a precise amount of pressurized fluid to the particular hydraulically-actuated device  113 . For instance, if hydraulically-actuated device  113  were a fuel injector, the amount of fuel injected can be about equal to the amount of fuel pressurized by unit pump  116 , thereby avoiding wasted energy that occurs by pressurizing fluid only to spill a substantial amount of that pressurized fluid back for repressurization because it is not needed for a particular injection event. Those skilled in the art will appreciate that although the preferred version of the present invention includes sleeves that open and close a spill port on a pumping piston, some other suitable structure could be substituted that accomplishes the same task, such as some other component that opens and closes the spill port incorporated into the piston for a portion of its reciprocation distance. 
     The above description is intended for illustrative purposes only, and is not intended to limit the scope of the present invention in any way. For instance, other types of control valves could be substituted for the example illustrated control valve without departing from the intended scope of the present invention. Thus, those skilled in the art will appreciate that various modifications can be made to the illustrated embodiment without departing from the spirit and scope of the present invention, which is defined in terms of the claims set forth below.