Patent Document

This application is 371 of PCT/US01/05413 Feb. 20, 2001 which claims benefit of No. 60/183,375 Feb. 18, 2000. 
    
    
     TECHNICAL FIELD 
     This invention relates to a variable delivery fluid pump and an electro-hydraulic control circuit therefor, and more particularly, to a fluid pump for use with a fuel injection system or other hydraulic system for an internal combustion engine. 
     BACKGROUND 
     In a common rail fuel injection system, high pressure fluid is supplied to the injectors from a high pressure fluid accumulator or manifold, which is referred to as a rail. To permit variation of the fluid pressure supplied to injectors from the rail, it is desirable to vary the delivery of fluid to the rail from one or more fluid supply pumps. Known common rail systems typically rely on either a single fluid pump that supplies fluid to the rail or a plurality of smaller displacement pumps that each supplies fluid to the rail. The volume and rate of fluid delivery to the rail has been varied in the past by providing a rail pressure control valve that spills a portion of the delivery from a fixed delivery pump to maintain the desired rail pressure. Both high pressure and low pressure common rail systems are known in the art. In high pressure common rail systems, high pressure fuel is supplied from the rail to electrically-controlled injection nozzles. In a low pressure common rail, an actuation fluid such as fuel or engine lube oil is supplied from the rail to unit injectors, whereby the actuation fluid is used to drive a fuel pressurization plunger that pressurizes the fuel to injection pressure prior to or during each injection event. 
     Variable delivery pumps are well known in the art and are typically more efficient for common rail fuel systems than a fixed delivery actuation fluid pump, since only the volume of fluid need to attain the desired rail pressure must be pressurized. For example, variable delivery has been achieved from an axial piston pump, e.g. a pump wherein one or more pistons are reciprocated by rotation of an angled swash plate, by varying the angle of the swash plate and thus varying the displacement of the pump. In such a pump, the swash plate is referred to as a “wobble plate”. Variable delivery has also been achieved in fixed displacement, axial piston pumps by a technique known as sleeve metering, in which each piston is provided with a vent port that is selectively closed by a sleeve during part of the piston stroke to vary the effective pumping portion of the piston stroke. An example of such a sleeve-metering pump is illustrated in commonly-owned U.S. Pat. No. 6,035,828. 
     While known variable delivery pumps are suitable for many purposes, known design are not always well suited for use with modem common rail fuel systems, which require fluid delivery to the rail to be varied with high precision and with rapid response times measured in microseconds. In addition, known variable delivery pumps are typically complex, may be costly, and are subject to mechanical failure. Moreover, in some known pumps such as the pump shown in commonly-owned U.S. Pat. No. 6,035,828, the relative positioning of the pumping pistons and the metering sleeves is controlled by way of an electro-hydraulic control circuit which receives high pressure fluid directly from the delivery gallery of the pump at high pressure and selectively spills that control fluid via an electrically-operable control valve. While pumps such as the one illustrated in U.S. Pat. No. 6,035,828, have performed well, room for improvement exists due the current need for small, high-precision passages and valve elements in the prior art as a result of the high fluid pressures present in the control circuit. 
     This invention is directed toward overcoming one or more of the problems described above. 
     SUMMARY OF THE INVENTION 
     In one aspect of this invention, a hydraulic pump system comprises a variable delivery, sleeve-metered pump having a plurality of pumping pistons and associated metering sleeves. The pumping pistons deliver pressurized fluid to a high pressure area at a pressure at least equal to a first pressure. An electrically-operated, hydraulic control circuit is operable to control the delivery of pressurized fluid from said pump by controlling the relative position between the pistons and their associated metering sleeves. The control circuit is in fluid communication with the high pressure area and has a pressure reducer to reduce pressure of fluid entering the control circuit to a control circuit pressure less than the first pressure. 
     In another aspect of this invention, the control circuit comprises a pressure reducing valve having an inlet in fluid communication with a high pressure area of the pump and having a valve outlet. The pressure reducing valve reduces the pressure of control fluid entering the control circuit to a predetermined control circuit pressure. A movable control member has a first control surface and a second control surface opposed with the first control surface, movement of the control member changing the relative positioning between the pumping pistons and their associated sleeves. A control line is in fluid communication with the pressure reducing valve outlet and has a first, relatively unrestricted passageway through which fluid pressure is applied to the first control surface and a second, relatively-restricted passageway through which fluid pressure is applied to the second control surface. An electrically operated control valve is fluidly connected with the control line to selectively control the relative fluid pressures applied to the first and second control surfaces. 
     In yet another aspect of this invention, the control circuit comprises a pressure reducing valve having an inlet in fluid communication with a high pressure area of the pump and having a valve outlet. The pressure reducing valve reduces the pressure of control fluid entering the valve to a predetermined control circuit pressure. A movable control member has a control surface, and movement of the control member changes the relative positioning between the pumping pistons and their associated sleeves. A control line is in fluid communication with the pressure reducing valve outlet and has a restricted passageway through which fluid pressure is applied to the control surface. A bias member applies force to the control member in a direction opposite the fluid pressure applied to the control surface. An electrically operated control valve is fluidly connected with the control line to selectively control the fluid pressure applied to the control surface. 
     In still another aspect of this invention, a method of controlling the delivery of pressurized fluid from a variable delivery, sleeve-metered pump is provided. The pump comprises a plurality of pumping piston and associated metering sleeves. The method comprising reciprocating the pistons to thereby deliver pressurized fluid to a high pressure area of the pump at pressure at least equal to a first pressure, delivering a portion of the pressurized fluid to a control circuit operable to selectively control the relative position between the pistons and their associated metering sleeves, reducing the pressure of the fluid delivered to the control circuit to a pressure less than the first pressure, and using the reduced-pressure fluid to control the relative position between the pistons and their associated metering sleeves, thereby controlling the delivery of pressurized fluid from the pump. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     FIG. 1 is a diagrammatic representation of a low pressure common rail fuel injection system in accordance with this invention. 
     FIG. 2 is an end elevation of a pump in accordance with this invention used in the fuel injection system shown in FIG.  1 . 
     FIGS. 3 through 5 are cross sections of the pump shown in FIG. 2 taken along lines  3 — 3 ,  4 — 4 , and  5 — 5  thereof, respectively. 
     FIGS. 6 and 7 are enlarged views of portions of FIGS. 4 and 5, respectively. 
     FIG. 8 is a diagrammatic illustration of a control circuit according to one embodiment of this invention. 
     FIG. 9 is a diagrammatic illustration of a control circuit according to a second embodiment of this invention. 
     FIG. 10 is a graph illustrating the relationship between control current I and pump output Q for a pump using the control circuit shown in FIG.  8 . 
     FIG. 11 is a graph illustrating the relationship between control current I and pump output Q for a pump using the control circuit shown in FIG.  9 . 
    
    
     DETAILED DESCRIPTION 
     FIG. 1 diagrammatically illustrates a fluid actuated diesel fuel injection system  10  with which this invention may be used. In particular, the fuel injection system includes a plurality of fluid-actuated injectors  12 , which may be unit injectors as illustrated or unit pumps injectors (not shown), powered via a variable delivery, fixed displacement fluid pump  14  in accordance with this invention. Actuation fluid is supplied to the pump  14  via an inlet  16 . High-pressure actuation fluid is supplied from the pump  14  to the unit pump injectors  12  via a manifold or common rail  18 . A conventional fuel transfer pump  20  supplies fuel to the injectors  12  via a common fuel rail  22 . The fuel system  10  illustrated in FIG. 1 preferably includes HEUI™ fuel injector available from Caterpillar Inc, preferably having a nozzle check valve operable independent of injection pressure, such as the injectors described in commonly-owned U.S. Pat. Nos. 5,463,996, 5,669,335, 5,673,669, 5,687,693, 5,697,342, and 5,738,075. 
     Of course, one skilled in the art will recognize that the injectors  12  may be hydraulically actuated fuel injectors having other configurations, such as those illustrated in patents granted to Sturman Industries and/or Oded E. Sturman (for example, U.S. Pat. No. 5,460,329) or otherwise using one or more high speed spool valves. Similarly, the pump  14  according to this invention may be used with conventional high pressure common rail systems or with the amplifier piston common rail system (APCRS) illustrated in the paper “Heavy Duty Diesel Engines—The Potential of Injection Rate Shaping for Optimizing Emissions and Fuel Consumption”, presented by Messrs. Bemd Mahr, Manfred Dürnholz, Wilhelm Polach, and Hermann Grieshaber; Robert Bosch GmbH, Stuttgart, Germany, at the 21 st  International Engine Symposium, May 4-5, 2000, Vienna, Austria. The pump  14  in accordance with this invention may also be suitable for use with fuels other than diesel fuel, such gasoline for example in a gasoline direction injection (GDI) application 
     With reference to FIGS. 2 through 7, the actuation fluid pump  14  is generally an axial, swash plate-type piston pump. The pump comprises an integral housing and barrel  24  that defines a plurality of cylinders  30  therein. Each cylinder  30  has slidably received therein a portion of a piston  32 , and a spring  34  is trapped between each piston  32  and the base of its corresponding cylinder  30 . Each piston  32  is connected at one end by a spherical mounting arrangement to a fixed angle swash plate  36 . More particularly, each piston  32  includes a spherical head  38  received within socket in a shoe  40  slidably mounted to the swash plate  36  by a hydrostatic bearing. As the swash plate  36  rotates, the pistons  32  are caused to move through a reciprocal stroke within the cylinders  30 . 
     The cylinders  30  and the pistons  32  cooperate to define a plurality of variable volume fluid compression chambers  42 . Each fluid compression chamber  42  has a delivery outlet  44  that is closed during the intake stroke by a conventional, but preferably cartridge-type, spring-biased check valve  46 . Each fluid compression chamber  42  also has a fluid inlet  48  to allow fluid to be drawn into the chamber  42  during the intake stroke. The fluid inlet  48  is preferably an inlet slot in the swash plate  36  that opens to ports in the heads  38  of the pistons  32 . The delivery outlets  44  each open to a common delivery gallery  50  in fluid communication with the outlet  52  of the pump. 
     Each fluid compression chamber  42  has a vent port  54  opening therefrom. As apparent, the vent ports  54  are operable to vent fluid from the fluid compression chambers  42  during a portion of the reciprocal stroke of the piston  32 . Each piston  32  has associated therewith a concentric sleeve  56  that is positioned to close the vent port  54  therein during portion of the piston stroke. The relative position of the sleeves  56  determines the effective pumping stokes of the pistons  32  and thus the output pressure of the pump. To provide a compact structure, the sleeves  56  are connected via a linkage  57  with a control shaft or member  58  located centrally between the pistons  32  and extending parallel to their axes of reciprocation. 
     The pump  14  also include a pilot control stage or control circuit, generally designated  60 , that is used to control axial movement of the control shaft  58  and thus control the position of the sleeves  56 . FIG. 8 illustrates diagrammatically the control circuit  60  shown in FIGS. 2 through 7. 
     With reference to FIGS. 2 through 8, high pressure oil from the pump delivery gallery  50  (or alternatively the pump outlet  52  or another high pressure area) is directed through a hydraulic passage  62  that leads to a conventional spool-type or other suitable pressure reducing valve, generally designated  64 , which is well known in the art and not described in detail herein. The valve  64  reduces the oil pressure in the control circuit  60  to a predetermined control circuit pressure significantly less than the maximum pump outlet pressure. For example, for pumps having a maximum outlet pressure on the order of 28-30 MPa, it is desirable to reduce the pressure in the control circuit  60  to around 4 MPa. The reduced pressure oil from the reducing valve  64  flows through a relatively-unrestricted passageway  65  and acts on a first control surface  66  forming part of or connected to the control shaft  58 . The oil also passes through a relatively-restricted passageway or control orifice  68  that creates a pressure differential whereby lower pressure oil acts on a second control surface  70  that is opposed to the first control surface  66 . The pressure differential between the first and second control surfaces  66 ,  70  creates a force imbalance that moves the control shaft  58  to the right. A spring  72  provides a force to move the control shaft  58  to the left. The direction of motion of the control shaft  58  is determined by the larger of the resultant fluid pressure force or the spring force. The control circuit  60  includes a control valve, generally designated at  74 , that is used to change the amount of oil that flows through the control orifice  68 . The control valve  74  comprises a solenoid or piezo actuator  76  that moves a pin  78  that is in contact with a conventional ball valve  80 . Of course, a poppet or spool valve could also be used. By varying the current to the valve actuator, the position of the ball valve  80  is varied, thus varying the amount of oil that is allowed to flow around the ball valve  80 . As the amount of oil flowing through the control valve  74  changes, the force imbalance on the control shaft  58  changes to control the motion of the control shaft  58 . In short, the specific current applied to the solenoid or piezo actuator  76  determines the amount of oil that flows through the control orifice  68 , which in turn determines the force differential on the control shaft  58 , which in turn determines the effective displacement of pistons  32 , which in turn determines the pump output. FIG. 10 illustrates, diagrammatically, the relationship between the current I that is applied to the control valve  74  and the output Q of the pump. 
     With reference now to FIGS. 9 and 11, an alternative embodiment  160  of a control circuit is shown diagrammatically. High pressure oil from the pump delivery gallery  50  is directed through a hydraulic passage  162  that leads to a conventional spool-type or other suitable pressure reducing valve, generally designated  164 . The valve  164  reduces the oil pressure in the control circuit  160  to a predetermined control circuit pressure significantly less than the maximum pump outlet pressure. The reduced-pressure oil also passes from a control line  165  through a relatively-restricted passageway or control orifice  168  that acts to reduce the fluid pressure from the predetermined pressure set by the reducing valve  164 . The oil then acts on a control surface  166  on the control shaft  58 . A force from spring  172  is applied opposite to the fluid force applied to control surface  166 . The force differential between the force applied to the control surface  166  and the spring force creates a force imbalance that moves the control shaft  58 . The direction of motion of the control shaft  58  is determined by the larger of the fluid pressure force applied to control surface  166  or the spring force. The control circuit  60  includes a control valve, generally designated  174 , that is used to change the amount of oil that flows through the control orifice  168 . By varying the current to the valve actuator, the amount of oil that is allowed to flow through the control valve  174  changes. As the amount of oil flowing through the control valve  174  changes, the force imbalance on the control shaft  58  changes to control the motion of the control shaft  58 . FIG. 11 illustrates, diagrammatically, the relationship between the current I that is applied to the control valve  174  and the output Q of the pump. 
     INDUSTRIAL APPLICABILITY 
     Prior pump designs of similar sleeve-metering configuration use full pump pressure to move the control shaft, and as a consequence, require a very small ball valve to allow only a small flow through the control valve. 
     Because the present designs relies on a reduced pressure, a larger ball valve can be used, which eases manufacture and improves pump control. Moreover, the pump can be operated using displacement control, for which there is a single pump output associated with each current level applied to the solenoid or piezo actuator. Thus, if a rail pressure change is needed, the current corresponding to the desired pressure is sent to the solenoid or piezo actuator to directly set the rail pressure that corresponds to the displacement set by the applied current. This is compared to prior designs, which are not admitted to be prior art, that utilize pressure control by sensing pressure in the rail and adjusting the sleeve position until the desired pressure is sensed in the rail. The pump configuration according to this invention also provides a compact and efficient package, in part as a result of the central positioning of the control shaft  58  and the end attachment of the control valve  60 . 
     This invention is illustrated in the context of a sleeve-metered pump is which the metering sleeves are movable relative to the pumping piston. However, one skilled in the art will recognize that this invention is also applicable to other pump configurations, including a pump configuration such as that illustrated in commonly-owned laid-open German patent application 199 60 569.6, filed on Dec. 15, 1999, which illustrates a pump in which the relative positioning of the pumping pistons with the “metering sleeves” is controlled by moving the pump swash plate with respect to the “metering sleeves”. In addition, while this invention is illustrated in connection with a fuel injection system, those skilled in the art will recognize that this invention is equally applicable to use with other hydraulic engine systems, such as engine valve actuators and/or compression release retarders. 
     Although the presently preferred embodiments of this invention have been described, it will be understood that within the purview of the invention various changes may be made within the scope of the following claims.

Technology Category: 2