Patent Publication Number: US-6901911-B2

Title: Pump and hydraulic system with low pressure priming and over pressurization avoidance features

Description:
TECHNICAL FIELD 
   The present invention relates generally to hydraulically-actuated systems used with internal combustion engines, and more particularly to a pump and hydraulic system with electronic control and biasing features for priming and prevention of over pressurization. 
   BACKGROUND 
   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. For example, variable angle swash plate type pumps are relatively complex, and thus are more prone to mechanical break down relative to simple fixed displacement type pumps. In addition, 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 order to control rail pressure. The Wear et al. system primes itself by having its pump biased to produce substantial output, even when system pressures are low, such as during a cold start. The Wear et al. pressure regulating valve and/or a separate pressure relief valve provide the means by which system over pressurization is avoided. 
   The present invention is directed to overcoming problems associated with, and improving upon, hydraulic systems. 
   SUMMARY OF THE INVENTION 
   In one aspect, a liquid pump includes a pump body with an outlet area and an inlet area disposed therein. At least one pump piston is moveably positioned in the pump body. An electro-hydraulic controller is attached to the pump body and is moveable between a first position at which the pump piston displaces fluid in a large proportion to the outlet area relative to the inlet area, and a second position at which the pump piston displaces fluid in a small proportion to the outlet area relative to the inlet area. A mechanical biaser is operable to bias the electro-hydraulic controller toward the first position, but a biasing hydraulic surface is oriented in opposition to the mechanical biaser for hydraulically biasing toward the second position, when available control pressure is high. A control hydraulic surface is oriented in opposition to the biasing hydraulic surface. 
   In another aspect, a method of operating a liquid pump includes a step of biasing a controller of the liquid pump with a mechanical biaser toward a high output position when a pressure differential between an outlet area and an inlet area of the liquid pump is low. The mechanical bias is overcome with a hydraulic biaser to bias the controller of the liquid pump toward a low output position when the pressure differential is high. 
   In still another aspect, a hydraulic system includes a source of fluid and a common rail with at least one hydraulic device fluidly connected thereto. An electro-hydraulically controlled liquid pump has an inlet fluidly connected to the source of fluid, and an outlet fluidly connected to the common rail. The liquid pump is biased to displace a relatively small amount of fluid toward the common rail when a pressure differential between the common rail and the source of fluid is large. The liquid pump is biased to displace a relatively large amount of fluid toward the common rail when the pressure differential is small. 

   
     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 a unit pump embodiment of the present invention; and 
       FIGS. 8   a-c  are graphs of sleeve position, pump pressure and controller electric current verses time for one example pump priming event. 
   

   DETAILED DESCRIPTION 
   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 engine brakes, could be substituted for, or added to, 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 provide 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 . Control valve  17  is preferably a portion of an electro-hydraulic controller  65 . 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 . Liquid 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. Nevertheless, those skilled in the art will appreciate that shaft  31  could be rotated indirectly by the engine or by some other machinery. A fixed angle swash plate  33  is attached to shaft  31 , but the invention also contemplates variable angle swash plates. 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 outlet area  40  past check valve  37  or spilled back into a low pressure inlet area  36  via a spill port  35 . 
   The proportion of fluid displaced by pistons  32  to the respective high pressure are  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 shaft  31  to a position in which virtually all the fluid displaced by the individual pistons  32  is displaced into high pressure space  40 , since spill ports  35  remain closed during the entire pumping stroke. The mechanical bias provided by spring  61  helps facilitate priming of pump  16 . Although electro-hydraulic controller  65  includes internal hydraulic surfaces that facilitate operation and control of output from pump  16  when system pressures are relatively high, these surfaces are of little help when starting the system at low pressure. Thus, spring  61  serves as a means by which the system can prime and come up to pressure during a cold start without reliance upon some stored source of pressurized fluid or some other means, in order to bias the electro-hydraulic controller  65  to a position that produces maximum output into high pressure space  40 . Those skilled in the art will appreciate that the pressure differential between high pressure space  40  and low pressure space  36  during a cold start up is small to non existent. 
   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 . Although the present invention prefers that electro-hydraulic controller  65  utilize sleeves that are moveable axially with respect to pistons  32  as a means by which spillage back to low pressure area  36  is controlled, those skilled in the art will appreciate that other spill control mechanisms could be substituted without departing from the intended scope of the present invention. 
   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 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 and displace a fixed volume of fluid. 
   Referring now to  FIG. 5 , the internal structure of electro-hydraulic controller  65 , which includes control valve  17  and sleeves  51 , is illustrated. Electro-hydraulic controller  65  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 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 . Passage  64  is preferably fluidly connected to low pressure area  36  via a passage that is not shown. When current is lower, fluid pressure in control volume  60  pushes on tip hydraulic surface  75  of poppet valve member  73 , causing it and armature  71  to move toward the right to open some fluid communication between control volume  60  and low pressure area  63  past valve seat  62 . Thus, depending upon the fluid pressure in control chamber  60  and the current supplied to solenoid coil  70 , the flow area past valve seat  62  can be precisely controlled. This in turn provides a means by which pressure in control volume  60  can be controlled to some pressure that is between that existing in the high pressure outlet area  40  and the low pressure inlet area  36 . 
   As stated earlier, actuator shaft  50  is normally biased away from coil  74  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 biasing hydraulic surface  53  that is exposed to fluid pressure in a biasing volume  53   a , which is always in fluid communication with the high pressure area  40  within pump  16  via a high pressure conduit  54 . Thus, biasing hydraulic surface  53  is oriented in opposition to spring  61  such that a hydraulic force would tend to bias shaft  50  toward a low output position as shown in FIG.  5 . This high fluid pressure in conduit  54  is channeled via central restricted communication passage  55  into control volume  60 . Fluid pressure in control volume  60  acts on a control hydraulic 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 to the right 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 hydraulically biased to a position as shown in  FIG. 5  in which no high pressure output is produced. Thus, when system pressure is relatively high, and current to solenoid coil  74  ceases, the pressure in control volume  60  acts upon tip hydraulic surface  75  of valve member  73  pushing it to an open position, which relieves pressure in control volume  60 . This lowered pressure force on control hydraulic surface  56  combined with the spring force produced by biasing spring  61  is preferably overcome by the biasing force on biasing hydraulic surface  53  such that shaft  50  will move toward coil  74  to a zero output position as shown in FIG.  5 . This aspect of the pump prevents over pressurization. 
   When pressure is low throughout the system, such as during a cold start, pressures everywhere in the pump are relatively low. When this occurs, biasing spring  61  provides a dominate force in electro-hydraulic controller  65  causing it to move away from coil  74  to a position as shown in  FIG. 2  in which substantially all of the fluid displaced by pump pistons  32  is pushed in the high pressure area. Thus, the pump includes a mechanical bias that facilitates priming, but that mechanical bias can be overcome at system pressures to bias the pump toward a low output position to prevent over pressurization in the event of electrical failure to electro-hydraulic controller  65 . 
   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. 
   The embodiment shown in  FIG. 7  is substantially similar to the earlier embodiment except that it is a unit pump containing only one pump piston verses a multi-piston swash plate type pump of the type earlier described. Nevertheless, it includes sleeve metering and an electro-hydraulic controller  165  similar in construction to that described earlier. In other words, a spring  161  normally biases sleeve  151  toward a position that produces maximum output in order to facilitate priming. Electro-hydraulic controller  165  also includes a biasing hydraulic surface  153  that is oriented in opposition to spring  161 . In addition, a control hydraulic surface  156  is oriented in opposition to biasing hydraulic surface  153 . Control valve  117 , which is a portion of electro-hydraulic controller  165 , controls the pressure force on control hydraulic surface  156  via high pressure fluid supplied from high pressure area  140  via high pressure control line  154 . The pressure on biasing hydraulic surface  153  is always relatively high. This embodiment also could differ from the earlier embodiment by the inclusion of a pressure reduction valve  155  so that the control function of the pump can consume less hydraulic fluid to perform its function. This aspect of the invention can be facilitated by appropriately sizing hydraulic surfaces  153  and  156  relative to spring strength  161  and other known factors. Thus, the earlier embodiment could also utilize a pressure reduction valve with appropriate spring strength and hydraulic surface area sizing to allow it to perform its control function with a reduced consumption of the pump&#39;s high pressure output. 
   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 electro-hydraulic controller  65 . The graphs of  FIGS. 6   a-d  reflect system operation when the pressure differential between outlet area  40  and inlet area  36  is high, such as during normal operation. When solenoid current is low, rail pressure remains at the lower end of its high pressure range. 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  (a controlled leakage flow area) 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 biasing hydraulic surface  53  is balanced against the combined forces from spring  61  and the hydraulic force acting on control hydraulic 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 to the position shown in  FIG. 2  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 toward the position shown in  FIG. 5  where the pistons have no effective pumping stroke. 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 . In addition, some steady drop in pressure will occur due to flow of high pressure fluid into control volume  60  and back to low pressure area  36  to perform the control function. 
   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. 
   Referring now to  FIGS. 8   a - 8   c , an example priming sequence for the pump of the present invention is illustrated. At the beginning time, the sleeve position is biased to a maximum output position by the mechanical biasing spring  61 ,  161 ; pressure throughout the pump is low; and current to the electro-hydraulic controller  65 ,  165  is at zero. As a pump piston(s) starts to move via rotation of its shaft as shown in  FIG. 2  embodiment or by the cam of the  FIG. 7  embodiment, fluid begins to be displaced into the high pressure area of the pump. This causes pressure in the outlet area to rise while pressure in the inlet area remains low. If no current were supplied to electro-hydraulic controller  65 , the pump would seek out an equilibrium pressure (EP) that reflects a balance between substantially all of the high pressure output of the pump being consumed through electro-hydraulic controller  65 . Thus, without any electrical current, the pump will come up to an operational pressure (EP) that produces sufficient pressure that the electro-hydraulic controller can operate effectively. This pressure is preferably high enough that the hydraulic system can still operate in a lower performance mode in the event of a voltage drop in the entire system. In other words, this pressure is preferably high enough to provide a limp home pressure that would allow the hydraulic system to operate. After reaching this equilibrium pressure, electric current can be supplied to electro-hydraulic controller  65 ,  165  to move the sleeves toward there maximum output position to raise pump outlet pressure to regular system levels. Once reaching the system pressure levels, if current to the electro-hydraulic controller  65 ,  165  is dropped back to zero, the sleeves will quickly move the their minimum or no output position as shown in  FIG. 8   a , and pressure will decay due to fluid leakage losses through electro-hydraulic controller  65 ,  165 . If current were not resupplied to adjust the pressure to some desired level, the pump would again seek out the equilibrium pressure level EP after some time delay. Thus, the present invention has a mechanical biasing feature that facilitates priming without any electrical current or stored fluid pressure, yet retains over pressurization prevention features via hydraulic biasing that prevents the system from becoming over pressurized when pressure is high and current to electro-hydraulic controller  65 ,  165  is disrupted for whatever reason. 
   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 very little 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. 
   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.