Patent Publication Number: US-6668801-B2

Title: Suction controlled pump for HEUI systems

Description:
This invention is a continuation of Ser. No. 10/123,887 filed Apr. 16, 2002, now abandoned entitled “Suction Controlled Pump for HEUI Systems” hereby incorporated herein by reference in its entirety, which is a continuation of Ser. No. 09/849,636 filed May 4, 2001, now U.S. Pat. No. 6,439,199 entitled “Pilot Operated Throttling Valve for Constant Flow Pump” hereby incorporated herein by reference in its entirety, which is a continuation-in-part of Ser. No. 09/553,285, filed Apr. 20, 2000, now U.S. Pat. No. 6,227,167 (“the &#39;167 patent”) issued on May 8, 2001, also incorporated herein by reference in its entirety. 
    
    
     This invention relates generally to multiple piston pumps and more particularly to a high pressure pump used in a hydraulically actuated electronically controlled unit injector (HEUI) fuel control system. The invention is particularly applicable to and will be described with specific reference to a constant flow, fixed displacement pump and the integration of the fixed displacement pump into a HEUI system. However, those skilled in the art will appreciate that the invention may have broader application and may be integrated into other hydraulic pump driven systems, such as vehicular steering systems. 
     This invention also relates to a control system for a fixed displacement, constant flow pump and more particularly to a hydraulically actuated electronically controlled unit injector (HEUI) fuel control system using the fixed displacement constant flow pump. The invention is particularly applicable to and will be described with specific reference to a throttling valve controlling metering of low pressure fluid into a high pressure pump used in a HEUI flow control system. However, the invention has broader application and may be applied to other systems using a constant flow, fixed displacement pump requiring fast response over a wide range of operating conditions such as vehicular steering systems as mentioned above. 
     BACKGROUND 
     A) Conventional Systems 
     As is well known, a hydraulically-actuated electronically-controlled unit injector fuel system has a plurality of injectors, each of which, when actuated, meters a quantity of fuel into a combustion chamber in the cylinder head of the engine. Actuation of each injector is accomplished through valving of high pressure hydraulic fluid within the injector under the control of the vehicle&#39;s microprocessor based electronic control module (ECM). 
     Generally, sensors on the vehicle impart engine information to the ECM  25  which develops actuator signals controlling a solenoid on the injector and the flow of hydraulic fluid to the injector. The solenoid actuates pressure balanced poppet valves such as shown in U.S. Pat. Nos. 5,191,867 and 5,515,829 (incorporated by reference herein). The poppet valves in the injector port high pressure fluid to an intensifier piston which causes injection of the fuel at very high pressures. The pressure at which the injector injects the fuel is a function of the hydraulic fluid flow supplied the injector by a high pressure pump while the timing of the injector is controlled by the solenoid. Both functions are controlled by the ECM to cause precise pulse metering of the fuel at desired air/fuel ratios to meet emission standards and achieve desired engine performance. Tightening emission standards and a demand for better engine performance have resulted in continued refinement of the control techniques for the injector. Generally the pump flow output has to be variable throughout the operating range of the engine. For example, one manufacturer may desire a constant pump flow throughout an operating engine speed range except at the higher operating engine speeds whereat the injectors are valving so quickly reduced pump flow may be desired even though more fuel is being injected by the injectors to the combustion chambers. Other manufacturers may desire to rapidly change pump flow at any given instant for emission control purposes. For example, the ECM may sense a step load change on the engine and impose a change in the fuel/air ratio to overcome the effects of a transient emission. Still further, the operating vehicular environment severely impacts oil viscosity affecting pump flow and injector performance. Viscosity of the hydraulic fluid is affected by several variables besides heat and is difficult to program into the ECM to fully account for its affect on system performance. 
     In a HEUI system, high pressure hydraulic actuating fluid is supplied to each injector by a high pressure pump in fluid communication with each injector through a manifold/rail fluid passage arrangement. The high pressure pump is charged by a low pressure pump. As noted in the &#39;867 patent, the high pressure pump is either a fixed displacement, axial piston pump or alternatively a variable displacement, axial piston pump. If a fixed displacement pump is used, a rail pressure control valve is required to variably control the pressure in the manifold rail by bleeding a portion of the flow from the high pressure pump to a return line connected to the engine&#39;s sump. For example, the &#39;867 patent mentions varying the output of the high pressure pump by the rail pressure control valve to pressures between 300 to 3,000 psi. A variable displacement pump can eliminate the rail control valve if the flow output of the variable pump can timely meet the response demands imposed by the HEUI system. The pumps under discussion are axial piston pumps in which the pump stroke (displacement) is determined by the angle of the swash plate. Variable displacement, axial piston pumps use various arrangements to change the swash plate angle and thus the piston stroke. Generally speaking, variable output, axial piston pumps do not have the reliability of a fixed displacement, axial piston pump and are more expensive. More significantly, the response time demands for pump output flow in a HEUI system is becoming increasingly quicker and a variable pump may be unable to change output flow within the time constraints of a HEUI system unless a rail pressure control valve is used. 
     A fixed displacement, high pressure pump is typically used in HEUI systems because of cost considerations. The pump is sized to match the system it is applied to. It is well known that the flow of a fixed displacement pump increases, generally linearly, with speed. Accordingly, the fixed displacement pump is sized to meet HEUI system demands at a minimal engine speed which is less than the normal operating speed ranges of the engine. Higher engine speeds produce excess pump flow which is dumped by the rail pressure control valve to return. The excess flow represents an unnecessary power or parasitic drain on the engine which the engine manufacturers have continuously tried to reduce. 
     For example, U.S. Pat. No. 5,957,111 shows a control scheme in which excess pump flow is passed to an idle injector but at a rate insufficient to actuate the injector. The system is stated to allow elimination of the rail pressure control valve and permit a more accurate sizing of the fixed displacement pump. However, the system does not avoid unnecessary parasitic engine power drains imposed by the pump. The pump must still be sized to produce a set flow sufficient to actuate the injectors at a low speed and that flow increases with pump speed. 
     B) The &#39;167 Patent 
     The &#39;167 patent discloses a fixed displacement, axial pump which in contrast to conventional axial piston pumps, eliminates the kidney shaped ports, rotates the cylinder, fixes the swash plate against rotation and establishes an orificed, suction slot inlet for each piston. The suction slot draws a constant volume of fluid into each pump cylinder once pump operating speed is reached to produce a constant flow output from the pump. The pump can therefore be designed to produce the maximum flow required by the HEUI system (i.e., at low operating speeds) which maximum does not increase when pump speed increases as in conventional fixed displacement pumps. The power otherwise expended to drive conventional fixed displacement pumps beyond their designed “maximum” is not required. Improved vehicle performance, better fuel consumption and decreased emissions results because the parasitic power drain is removed. 
     Additionally, and as noted above, there are times during the vehicle&#39;s operation where less flow from the required “maximum” is sufficient to operate the injectors and desired for better injector performance, enhanced fuel consumption, etc. In the prior applications, it was demonstrated that controlling the flow of fluid to the constant volume high pressure pump by a throttling valve could produce a constant pump output flow at any desired level. The results and benefits achieved by the constant flow pump as discussed above relative to the maximum output sizing consideration, can therefore be achieved throughout the operating range of the pump by a throttling valve at the pump inlet. Parasitic power drains on the system are thus alleviated over the entire operating range of the engine. 
     The throttling valve generally disclosed in the &#39;167 patent was simply a solenoid operated valve under the control of the ECM and similar to the high pressure, axial pressure control valve (RPCV) currently used in conventional systems. Because the solenoid valve is controlling the flow of a low pressure pump, its sizing is reduced decreasing its cost. While the solenoid operated valve can throttle the flow to the inlet of the constant flow pump, the viscosity changes in the hydraulic fluid such as the variations that can occur between ambient vehicular start-up temperatures and the sudden fluid flow changes occurring during normal operating conditions, such as that occurring during vehicle acceleration or deceleration, impose requirements on a conventional solenoid valve which are difficult to achieve. 
     SUMMARY OF THE INVENTION 
     It is therefore a principal object of the invention to provide a fixed displacement multiple piston pump which can be sized for a HEUI or other hydraulic system to alleviate or minimize engine power or parasitic drains imposed on the engine attributed to the associated bleeding of excess capacity pump flow. 
     This object along with other features of the invention is achieved by a constant flow, fixed displacement, piston pump which includes a non-rotatable cylinder containing a plurality of piston bores spaced about a centerline of the pump. A rotatable shaft having a formed shaft portion is journalled in the pump. Within each bore a piston is movable and has one end extending through a bore end and in contact with the formed shaft portion while the piston&#39;s opposite end is adjacent an outlet check valve at the opposite bore end. The pump has a discharge chamber in fluid communication with all piston outlet check valves and with the pump outlet. Each piston bore has suction slot of set area in fluid communication with the pump inlet which is sized as a function of timed flow through an orifice. The suction slot is transversely positioned at a set distance between the piston bore ends and sealed and opened by axial movement of each piston within its bore whereby fluid displaced into the piston bore decreases during the piston suction stroke in fixed relationship to increases in shaft rotational speed after the operating speed of the pump has been reached to produce a constant displacement pump throughout the operating range of the pump. 
     An important feature of the invention is achieved by an improvement to an internal combustion engine having a hydraulically actuated, electronically controlled fuel injection system of the type including a fuel injector valving high pressure fluid in response to commands from an ECM to timely inject a metered quantity of fuel to the engine&#39;s combustion chamber. The injector is in fluid communication with the outlet of the high pressure pump which in turn has an inlet in fluid communication with a low pressure pump. The improvement includes a fixed displacement high pressure pump, as described above, which produces a constant output flow of fluid at all operating speeds of the pump whereby the pump can be sized to match the flow demands of a HEUI system without placing excessive or unneeded power demands on the engine. 
     In accordance with another important aspect of the invention, the improved system includes the provision of a pressure control throttling valve at the inlet of the high pressure pump whereby the generally constant high pressure flow from the high pressure pump can be reduced to lower displacement flow values in response to commands from the ECM without placing any load on the engine to develop a pump pressure higher than what is required to actuate the HEUI system. 
     In accordance with another aspect of the invention, an annular discharge chamber is in fluid communication with the outlet check valve and the outlet port of the pump. The outlet check valve may be a reed flapper valve whereby high pressure fluid pumped by all cylinders in the pump is united in the discharge chamber to dissipate pump pulsations. 
     In accordance with a still further aspect of the invention, the high pressure pump has a housing defining a chamber therein and the cylinder is fixed to the housing which also journals the rotatable shaft therein. The housing also has an annular inlet chamber in fluid communication with the bore slots and a drain passage is provided for fluid communication between the housing chamber and the inlet chamber whereby internal pump leakage is drained through the pump inlet avoiding external pump drain lines when the pump operates in a hydraulic system where the pump inlet is not pressurized. 
     It is an object of the invention to provide a fixed displacement pump having generally constant output flow throughout its operating speeds. 
     It is a primary object of the invention to provide a fixed displacement pump for use in any vehicular hydraulic system driven by the vehicle&#39;s engine which reduces or minimizes the power drain imposed by the pump on the engine. 
     It is another object of the invention to provide a fixed displacement pump for use in a HEUI system which provides a constant flow of pressurized fluid over the operating range of the pump to allow a better and/or more consistent control of the injector over the operating range of the engine. 
     It is another object of the invention to provide a hydraulic circuit for actuating a hydraulically actuated electronically controlled fuel injector which delivers constant pump flow over an operating pump speed range with an ability to throttle the flow on demand while decreasing power demands of the pump on the engine. 
     Still yet another object of the invention is to provide a fixed displacement pump for use in a HEUI system which alleviates the need for a rail pressure control valve, or, alternatively, allows for use of a smaller, less expensive rail pressure control valve. 
     Still yet another object of the invention is to provide a fixed displacement pump which is able to provide fluid to a hydraulically actuated, electronically controlled fuel injector that simulates or improves upon the performance level achieved by a variable displacement pump. 
     Still yet another object of the invention is to provide an improved low cost high pressure pump for use in an HEUI system. 
     A still further general object of the invention is to provide a fixed displacement pump producing a constant flow of pressurized hydraulic fluid over an operating speed range of the pump for use in any number of vehicular hydraulic systems which use the power from the engine to control the hydraulic system. 
     These and other objects, features and advantages of the invention will become apparent to those skilled in the art upon reading and understanding the Detailed Description of the Invention set forth below. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     The invention may take form in certain parts and arrangement of parts, a preferred embodiment of which will be described in detail and illustrated in the accompanying drawings which form a part hereof and wherein: 
     FIG. 1 is a prior art schematic illustration of a HEUI fuel injection system; 
     FIG. 2 is a prior art schematic hydraulic actuating fluid circuit diagram for the injection system shown generally in FIG. 1; 
     FIG. 3 is a constructed graph of pump flow versus speed for a conventional fixed displacement pump and for the fixed displacement pump of the present invention; 
     FIG. 4 is a sectioned side elevation view of the fixed displacement pump used in the present invention; 
     FIG. 4A is a sectioned elevation view similar to that shown in FIG. 4 but through a section about 90 degrees to the pump section shown in FIG. 4; 
     FIG. 5 is a plan view of the reed flapper valve used in the pump; 
     FIG. 6 is an enlarged view of a portion of the piston bore seal of the pump of the present invention; 
     FIG. 7 is a constructed graph showing plots of pump flow, pressure and torque versus speed of the pump used in the present invention; 
     FIG. 8 is a partial sectioned view showing a modification to the suction slot and pump of the preferred embodiment; 
     FIG. 9 is a sectioned view showing a modification to the vent orifice of the pump; 
     FIG. 10 is a constructed graph showing various flow rates achieved by the pump of the present invention; 
     FIG. 11 is a schematic hydraulic circuit of the present invention similar to FIG. 2; 
     FIG. 12 is a schematic hydraulic circuit similar to FIG. 11 but schematically showing the components of the throttling valve of the present invention; 
     FIG. 13 is a sectioned view of the throttling valve of the present invention; 
     FIG. 14 is a perspective view of the sleeve used in the flow control valve of the present invention; 
     FIG. 15 is a sectioned view of a solenoid actuated pressure control valve used in the throttling valve of the present invention; and, 
     FIG. 16 is a schematic view of an alternative embodiment of the present invention similar to FIG.  12 . 
     Before one embodiment of the invention is explained in detail, it is to be understood that the invention is not limited in its application to the details of construction and the arrangements of the components set forth in the following description or illustrated in the drawings. The invention is capable of other embodiments and of being practiced or being carried out in various ways. Also, it is understood that the phraseology and terminology used herein is for the purpose of description and should not be regarded as limiting. The use of “including” and “comprising” and variations thereof herein is meant to encompass the items listed thereafter and equivalents thereof as well as additional items. The use of “consisting of” and variations thereof herein is meant to encompass only the items listed thereafter. The use of letters to identify elements of a method or process is simply for identification and is not meant to indicate that the elements should be performed in a particular order. 
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
     A) The HEUI System 
     Referring now to the drawings wherein the showings are for the purpose of illustrating a preferred embodiment of the invention only and not for the purpose of limiting the same, reference is first had to a description of a prior art HEUI system as shown in FIGS. 1 and 2 since the present invention may be perhaps best explained by reference to an existing arrangement. 
     The system shown in FIGS. 1 and 2 will only be described in general terms and reference should be had to the patents discussed in the Background for a more detailed explanation of the system including the operation of the fuel injector, per se, which is not shown in detail herein. 
     Referring first to prior art FIG. 1, there is diagrammatically shown an HEUI fuel injection system  10  which includes a plurality of unit fuel injectors  12 . A fuel pump  13  draws fuel from the vehicle&#39;s fuel tank  14  and conditions the fuel at a conditioning station  16  before pumping the fuel to individual injectors  12  as shown. One or more fuel return lines  17  is provided. The fuel supply system as shown is separate and apart from the hydraulic system which actuates fuel injectors  12 . It is understood that the engine fueled by injectors  12  is typically a diesel engine and that diesel fuel (fuel oil) can be optionally used as the fluid to power injectors  12 . In the preferred embodiment, engine oil is used to actuate injectors  12 . Those skilled in the art will recognize that the present invention is functional in those systems which use diesel fuel pumped under high pressure to actuate injectors  12 . 
     Fuel injectors  12  are actuated by hydraulic pressure which, in turn, is regulated by signals generated by an electronic control module, ECM  18 . ECM  18 , in response to a number of sensed variables, generates electrical control signals which are inputted at  19  to a solenoid valve in each fuel injector  12  and to a rail pressure control valve  20  which determines the pressure of engine oil pumped to fuel injectors  12  by a high pressure pump  32 . 
     More particularly, ECM  18  receives a number of input signals from sensors designated as S 1  through S 8 . The sensor signals represent any number of variables needed by ECM  18  to determine fueling of the engine. For example, input signals can include accelerator demand or position, manifold air flow, certain emissions sensed in the exhaust, i.e., HG, CO, NOx, temperature, engine load, engine speed, etc. In response to the input signals, ECM accesses maps stored in look-up tables and performs algorithms, also stored in memory, to generate a fueling signal on S 9  which is inputted as an electrical signal to rail pressure control valve  20  and a signal on S 10  which takes the form of an electrical signal actuating a solenoid in injector  12 . Injector  12  is entirely conventional and can take any one of a number of known forms. For purposes of this invention, it is believed sufficient to state that high pressure fluid from a high pressure pump is supplied to the injectors. The pump fluid, which is supplied to injectors  12  is, in the preferred embodiment, engine oil and drains from the injectors back to the engine sump (oil pan) through the engine&#39;s case (valve housing). Generally, pressure balanced poppet valves actuated by the solenoid, direct high pressure pump fluid against a pressure intensifier within injector  12 . The pressure intensifier pressurizes diesel fuel to very high pressures (as high as 20,000 psi while high pressure pump pressure is not higher than about 4,000 psi) and ejects a pulse of fuel at this high pressure into the engine&#39;s combustion chamber. Poppet valve design, the staging or sequencing of the poppet valves, the degree of solenoid actuation, etc. will vary from one engine manufacturer to the next to generate a particular fuel pulse matched to the ignition/combustion characteristics of the combustion chamber formed by the geometry of the engine&#39;s piston/cylinder head. Various pulses such as square, sine, skewed, etc. can be developed by the injector  12  in response to solenoid signals from ECM  18 . 
     As noted in the Background, the HEUI system has enjoyed its widespread acceptance because its operation is not affected by the speed or load placed on the engine. However, the HEUI system requires high pressure actuating fluid to operate and the flow rate of the fluid has to be variable on demand to produce the desired feed pulse from the injector. Again, how the pulse is developed is beyond the scope of this invention, it is sufficient for an understanding of the present invention to recognize that the pump supplying actuating fluid to the injectors must achieve a minimum flow rate which allows the injector to achieve maximum fuel pressure. Once the high pressure pump achieves this output, the HEUI system, through rail pressure control valve (RPCV)  20  may reduce the pump flow on demand for any number of reasons to produce a desired fuel pulse. For example, one engine manufacturer may desire a constant pump flow through the operating range except that at high operating engine speeds, the poppet valves within injectors  12  may cycle so quickly that it is desirable for pump flow to be reduced. That is the pressure of the fluid can be transferred instantaneously before the hydraulic fluid drain through the injector “catches up”. Another manufacturer may sense load changes imposed on the engine and throttle the high pressure pump flow, at any engine operating speed, for emission purposes. In conventional systems, high pressure pump  32  supplies excess flow to injectors  12  which excess flow is returned to drain through RPCV  20  and the excess flow continues to increase as the pump speed increases. While rail pressure control valve  20  has been refined to timely respond to ECM demands, it should be clear that if the pump&#39;s excess flow can be reduced to more closely model system flow demands, the size (and expense) of rail pressure control valve  20  can be reduced. 
     As shown in prior art FIGS. 1 and 2, oil from the vehicle&#39;s conventional oil pump or low pressure pump  23  is cooled by a conventional radiator core  26 . A low pressure oil stream produced by a pressure valve  28  fills a priming reservoir  30  which is in fluid communication with the inlet end of a high pressure pump  32 . High pressure pump  32  includes the components shown in FIG. 2 within dot-dash line indicative of pump housing  32   a.  High pressure pump  32  pressurizes the engine oil at the high pressure pump&#39;s outlet (now termed actuating oil) which is in fluid communication with common rail passage  33  in the manifold which, in turn, is in fluid communication with rail branch passages  34  leading to actuating ports within individual fuel injectors  12 . In the prior art arrangement shown in FIGS. 1 and 2, a vee-type engine is used so there are two manifolds and two sets of rails. Also, for convenience in notation, reference to “rail” means the common rail passage  33  and rail branch passages  34  and can optionally include the actuating oil supply line  35  leading from the outlet of high pressure pump  32  to the manifold. When high pressure pump  32  is operating, pressure of the actuating oil in manifold/rail passages  33 ,  34  as noted above is determined by the actuation of rail pressure control valve  20  which is backed up with a safety relief valve  21 . 
     Referring now to prior art FIG. 2, priming reservoir  30 , in addition to functioning as an oil reservoir supplying oil to the inlet of high pressure pump  32 , functions also as a reservoir to maintain oil in the high pressure pump inlet supply line  38  and oil in high pressure pump  32  as well as oil in the manifold/rail passages  33 ,  34  when high pressure pump  32  doesn&#39;t operate. This is achieved by physically positioning priming reservoir  30  at an elevation above the inlet port of high pressure pump  32  and above manifold/rail passages  33 ,  34  and specifically, the use of a stand pipe  37  at that elevation to establish a gravity flow from priming reservoir  30 . Make-up oil flows past a one way check valve  39  (oil ferry) through an optional flow restriction orifice  40  in a bypass line  41  which communicates with actuating supply line  35 . Orifice  40  in combination with check valves  36  also functions to control Helmholtz resonance for balancing pressure surges or waves between the two manifolds for the vee-type engine illustrated. The make-up oil from priming reservoir  30  thus flows to the actuating supply line  35  and then to manifold/rail passages  33 ,  34 . Make-up oil also flows through actuating supply line  35  to the outlet of high pressure pump  32 . Leakage within high pressure pump  32  returns to crank case sump  24  through a fluid leakage supply line  43 . When priming reservoir  30  is filled by low pressure pump  23  excess oil and air is vented for return to crank case sump  24 . In the prior art FIG. 2 this occurs through an overflow return line  44  which includes an orifice  45  to maintain a slight pressure in priming reservoir  30 . It is or should be clear that in the HEUI system embodiment shown in FIGS. 1 and 2, the inlet of high pressure pump  32  during engine operation is charged through reservoir  30  at the pressure of low pressure pump  23 . 
     This invention, in its broad sense, is not limited to a HEUI system. However, like the HEUI system disclosed in FIGS. 1 and 2, a source of fluid, at some low pressure, must be available to charge the inlet of the high pressure pump. 
     B) The High Pressure Pump 
     Referring now to FIG. 3, there is shown a constructed graph plotting pump speed along the x-axis and pump flow along the y-axis for a fixed displacement pump. As is well known, pump flow increases, generally linearly, as a function of pump speed for a fixed displacement pump as shown by the dotted trace  50 . For reasons which will be explained in detail below, pump  55  of the present invention operates as a conventional fixed displacement pump in the sense that increasing pump speed increases pump flow. However, in the present invention, when a pump critical speed, hereinafter termed “operating speed”, is reached, the pump flow is constant notwithstanding increases in pump rotational speed. The operating speed of pump  55  of the present-invention is shown by the solid line indicated by reference numeral  51 . Further, for reasons discussed below it is possible for the pump flow of pump  55  to be decreased at any operating pump speed and this is indicated by dot-dash line  52  in FIG.  3 . 
     Referring now to FIGS. 4 and 4A, high pressure fixed displacement axial piston pump  55  includes a pump body  56  which is sealing secured to an end body casting  57  to define a body chamber  58  extending along pump centerline  60 . Fixed to pump body  56  and end body casting  57  is a piston cylinder  62  containing a plurality of piston bores  63  circumferentially spaced about pump centerline  60 . Disposed and axially movable within each piston bore  63  is a piston  64 . 
     Journalled within body chamber  58 , as by a sleeve bushing  65 , is a gear driven shaft  66 . Shaft  66  is rotatably sealed within body chamber  58  by a shaft seal  68  at one end. A portion of shaft  66  is formed as a swash plate  70 , one end of which contacts a thrust bearing  72 . Alternatively, swash plate is affixed or keyed to shaft  66  so as to be rotatable therewith. A tail shaft  69 , longitudinally extending along centerline  60 , is received within a central opening  71  extending through piston cylinder  62  and seated against a central recess in end body casting  57 . Tail shaft  69  has a necked down stem portion  73  extending out of central opening  71  which receives a spherical bearing  74 . Spherical bearing  74  is biased by a spring  75  in a direction that pushes spherical bearing  74  off stem  73  and is retained in the assembled position shown in FIGS. 4 and 4A because it engages, at its spherical bearing surface, a central opening in a slipper retainer plate  76 . The circular central opening in slipper retainer plate  76  has a diameter less than the outside spherical diameter of spherical bearing  74 . Slipper retainer plate  76  has circumferentially spaced, radially outward openings that receive and maintain socket shaped slippers  78  in contact with swash plate  70  and each piston  64  has a ball end  80  received within the socket of an associated slipper  78 . Thus, pistons  64 , which are fixed (although longitudinally movable) vis-a-vis stationary piston cylinder  62 , likewise fix slippers  78  vis-a-vis the ball/socket connection which in turn fix the position of slipper retainer plate  76  and slipper retainer plate  76  prevents spherical bearing  74  from leaving stem portion  73  under the bias of spring  75 . Spring  75  thus maintains, through the connections described, slippers  78  in contact with swash plate  70  while slipper retainer plate  76  pivots or swivels about spherical bearing  74  upon rotation of swash plate  70  relative to piston cylinder  62 . Note that while tail shaft  69  is not rotated by gear driven shaft  66 , tail shaft  69  and the opening in spherical bearing  74  which receives stem portion  73  are cylindrical in the preferred embodiment. This may enhance the swivel/pivoting motion of slipper retainer plate  76  relative to spherical bearing  74 . Other arrangements can be employed to allow rotation of swash plate  70  relative to fixed piston cylinder  62  while maintaining a spring bias against spherical bearing  74 . However, the general arrangement of slipper retainer  76 /spherical bearing  74  with the spherical bearing spring biased to a set axial position by spring  75  centered on centerline  60  produces a stable arrangement allowing for smooth axial motion of pistons  64  throughout the speed ranges of pump  55 . Other arrangements use offset varying spring forces in the piston bore to maintain slipper/swash plate contact. 
     As described thus far, pump  55  is different from typical axial piston pumps in which the cylinder rotates relative to a stationary swash plate. In pump  55 , rotation of swash plate  70  causes piston  64  to axially move in bore  63  through spherical bearing  74 , retainer plate  76  and slippers  78 /piston ball end  80 . For definition, rearward (toward the left when viewing FIG. 4) movement of piston  64  out of bore  63  at the ball end  80  side of piston  64  is a “suction stroke” of piston  64  while forward (towards the right when viewing FIG. 4) movement of piston  64  into piston bore  63  produces a “compression stroke” of piston  64 . Movement of piston  64 , caused by relative rotation of swash plate  70  and piston  62 , is conventional, although typically swash plate  70  is stationary. 
     Adjacent the forward end  81  of piston  64 , a vent insert  86  is inserted at the discharge end of piston bore  63 . Vent insert  86  has a vent orifice  87  formed therein which communicates through a one-way check valve with an annular discharge chamber  88  formed in end body casting  57  which in turn is in fluid communication with a pressurized outlet port  90  of pump  55 . Unlike traditional axial piston pumps, there are no kidney shaped inlet and outlet passages in fluid communication with the piston bore vent orifice as the piston cylinder rotates to sequentially communicate the vent orifice with a kidney shaped inlet passage during the piston&#39;s suction stroke and with a kidney shaped outlet passage during the piston&#39;s compression stroke. In the traditional axial piston pump, when the piston bores rotate to switch from the inlet kidney shaped passage to the outlet kidney shaped passage, the bores pass over lands which produce or contribute to pulsation of the fluid, especially at high pump speeds. This is avoided or minimized in pump  55  by having all piston bores  63  communicate through a check valve with a common annular discharge chamber  88  which unites or unifies the flow from piston bore  63  during the compression stroke of piston  64  while the check valve prevents flow of fluid from annular chamber  88  into piston bore  63  during the suction stroke of piston  64 . While annular discharge chamber  88  could be a centrally positioned chamber and relatively large, preferably, it is ring shaped and in the nature of a passageway, as shown in FIG. 4, which has been found to produce consistent, somewhat non-pulsing flow through outlet port  90 . 
     As best shown in FIGS. 4 and 6, pump body  56  has an inlet passage  79  which is in fluid communication with an annular inlet chamber  83  in piston cylinder  62  that terminates at an orificing slot  84  that establishes an opening in piston bore  63 . In the preferred embodiment, slot  84  is opened for some travel distance of piston  64  during the suction stroke and closed during the compression stroke of the piston. In the preferred embodiment, hydraulic fluid at inlet passage  79  is at low pressure (about 20-60 psi) from low pressure pump  23 . Fluid flows through orificing slot  84  during the time slot  84  is opened establishing an orifice in fluid communication with piston bore  63 . As the speed of the pump increases, the time that slot  84  is opened during the suction stroke of piston  64  decreases. Accordingly, successively smaller quantities of fluid enter piston bore  64  during the suction stroke as pump speed increases to produce a constant flow of fluid from outlet port  90 . 
     Specifically, the variable output of pump  55  is achieved by sizing suction slot  84 . Flow is controlled through suction slot  84  by the orifice equation: 
     
       
         
           QA·ΔP 
           1/2 
           ·t 
         
       
     
     Where “Q” is the flow, i.e., the quantity of fluid flowed for a time through the slot, “A” is the area, “ΔP” is the pressure drop across the slot, and “t” is the time the slot is open. The maximum displacement is achieved when time is of a magnitude that causes no limitation on the flow, i.e., it is of sufficient duration to fill the piston bore volume. That is to say, for maximum pump displacement the only controlling factors are the size of the orifice and the pressure drop. Time is inversely proportional to pump speed and causes no limitation on flow up to a certain critical or “operating” pump speed. Beyond that critical or operating speed, the flow through slot  84  is limited causing a constant amount of flow regardless of speed. 
     In the preferred embodiment, slot  84  is positioned rearwardly in piston bore  63  as shown in FIGS. 4 and 6. However, other arrangements such as shown in FIG. 8 are possible. In FIG. 8, suction slot  84  is positioned forwardly in piston bore  63  and equipped with a ball check valve  85 . Slot  84  is thus open for a longer travel distance during the suction stroke of piston  64  than that shown in FIGS. 4 and 6. However, in accordance with the orifice equation above, the size of slot  84  is controlled to produce constant flow over the operating speed. Other slot arrangements will suggest themselves to those skilled in the art. Conceptually, suction slot  84  could be positioned rearward in piston bore  63  so that it is not uncovered by piston  64  and piston could have an orifice opening in its sidewall, fitted with a check valve, allowing fluid to pass through piston  64  to fill piston bore  63  during the suction stroke. All of these arrangements establish an orifice, of a preset size, which is in timed fluid communication with inlet fluid to vary the volume of fluid admitted to piston bore  63  as a function of pump speed. In contrast, axial piston pumps which do use a stationary swash plate maintain fluid communication with the inlet throughout the suction stroke by a feed arrangement which assures filling the piston bore with fluid. 
     In the embodiment of pump  55  illustrated in FIG. 4, forward end  81  of piston  64  is open and a bleed passage  92  formed in piston ball end  80  provides forced lubrication to slipper/swash plate contact surfaces. Optionally, if pump  55  is not charged with pressurized inlet fluid at inlet  79 , internal leakage within pump which collects in body chamber  58  can be routed back to drain through inlet  79  by the provision of an optional drain passage  89  providing fluid communication between body chamber  58  and inlet chamber  83 . Pump  55  may not be charged with pressurized inlet fluid in vehicular hydraulic steering applications. In the HEUI system described in FIGS. 1 and 2, pump inlet  79  is at low pressure and pump leakage occurs at front shaft seal  68  which is conventional. 
     As noted, output of fluid from all piston bores  63  is united or unified in annular discharge chamber  88  which has the effect of dampening pulsations attributed to any specific piston  63  during its pressure stroke. In order to prevent back flow of pressurized fluid into piston bores  63  having pistons in a suction stroke travel mode, a check valve is positioned at the outlet of vent orifice  87 . In the preferred embodiment, a reed type flapper valve  94 , best shown in FIGS. 5 and 6, is positioned at the outlet of vent orifice  87  and held in spaced relationship by a vent plate  95  as shown in detail in FIG.  6 . Flapper valve  94  closes when the pressure of the fluid in piston bore  63  is less than the pressure of the fluid in outlet chamber  88 . Flapper valve  94  opens when the pressure of the fluid within piston bore  63  equals or exceeds the pressure of the fluid in annular outlet chamber  88 . In the preferred embodiment, as shown in FIG. 5, pump  55  has nine piston bores  63  and the relative diameter of discharge chamber  88  is shown by dot-dash circle  93 . An alternative to reed flapper valve  94  is a check valve such as ball check valve  97  fitted into vent insert  86  as schematically illustrated in FIG.  9 . 
     Reference can now be had to FIG. 7 which is a constructed graph showing performance of the pump design of FIG.  4 . Pump pressure is shown as the trace passing through dot dash line indicated by reference numeral  98 . Pump torque is shown by the trace passing through dash line indicated by reference numeral  99  and pump flow is shown by the trace passing through solid line indicated by reference numeral  100  at various rotational speeds of shaft  66 . FIG. 7 was constructed with inlet pump pressure at one atmosphere and pump fluid at 120 degrees F. As pump speed increases, flow of fluid through suction slot  82  increases with increasing pump speed until a critical or operating speed of the pump is reached whereat a knee  101  is formed in flow curve  100 . In the graph of FIG. 7, the flow limiting critical or operating speed of the pump is shown to occur at about 900 rpm. As trace  100  shows, further increase in speed of the pump during this operating range does not result in fluid flow increases. As a matter of definition and as used herein and in the claims, “operating speed” of pump  55  means the speeds at which pump  55  generally produces constant output flow as shown, for example, by trace  100  after knee  101 . It should also be noted that torque curve  99  shows torque decreasing with increases in pump speed during the “operating speed” of pump  55 . Torque decreases due to the relationship between torque and effective displacement. That is, 
     
       
         
           TN·D 
         
       
     
     Where “T”=torque, “N”=speed and “D” is effective displacement. Effective displacement of fluid from each piston bore  63  decreases during the suction stroke as explained above. Further, for a constant inlet pressure producing a constant pressure drop, it is possible to control the start of the “operating speed” or knee simply by sizing only the slot area. 
     It is also possible to achieve secondary control of variable pump displacement output by controlling the pressure of the fluid at the inlet side of suction slot  82 . In the HEUI application, and as noted, low pressure pump typically delivers fluid at inlet  79  at about 20-60 psi. This affects flow through suction slot  82  by the orifice equation set forth above. Changing inlet pressure changes the pressure drop across the orifice and produces a different flow curve. This is best shown by reference to FIG. 10 which shows operating speed flow curves  102 A,  102 B and  102 C. Inlet pressure is constant for each curve but the inlet pressure for curve  102 A is less than that for inlet curve  102 B which is less than that for inlet curve  102 C. In each case, an operating speed is reached whereat constant pump flow occurs but knee  101  at which the pump transitions to its operating (or critical) speed shifts with increasing inlet pressure. FIG. 10 shows that it is possible, by throttling the inlet flow, to variably control the pump&#39;s output flow when the pump is within its operating speed range. That is, the output flow of pump  55  at any speed within the pump&#39;s operating speed can be controlled by throttling the inlet flow such as shown by curve portion  52  of FIG.  3 . Conceptually, placing RPCV  20  upstream of pump  55  can achieve the valving now achieved by RPCV  20  downstream of conventional high pressure pump  32  but without the parasitic power drain of a conventional high pressure pump  32 . 
     Referring now to FIG. 11, there is shown a portion of the hydraulic circuit shown in FIG. 2 of the prior art modified to incorporate the operating characteristics of pump  55 . Components illustrated in FIG. 11 which are functionally similar to the components illustrated and discussed above with respect to prior art FIGS. 1 and 2 will be assigned the same drawing reference numerals as that used in describing the prior art. More particularly, FIG. 11 is characterized by the addition of a solenoid operated throttling valve  105  functionally similar to RPCV  20  and actuated by ECM  18 . That is, ECM  18  knows the constant flow of axial piston pump and actuates throttling valve  105  to drop the constant flow to any lesser value. (A throttling valve port shown by reference numeral  106  in FIG. 4 is in fluid communication with inlet port  79 .) The constant flow value is set at minimum system flow requirements plus a safety factor required by the system. In the preferred embodiment, RPCV  20  is eliminated from FIG.  11 . It is shown in FIG. 11 because of a slight fractional second delay which can elapse from the time throttling valve  105  is actuated to the time the reduced flow appears at pump outlet  90 . Some manufacturers may desire a millisecond response so RPCV  20  is shown in FIG.  11 . In such instance, ECM has to co-ordinate throttling valve  105  and RPCV  20 . A downsized RPCV  20  would be employed and actuated, in theory, for a fractional second until pump output realized the setting of throttling valve  105 . Alternatively, RPCV  20  can be eliminated. 
     C) The Throttling Valve 
     As discussed above and illustrated in FIG. 11, the RPCV  20 , which was heretofore placed downstream of high pressure pump  55 , can be placed upstream of the high pressure pump to avoid the parasitic power drain of the conventional high pressure pump  32  (FIGS.  1  and  2 ). Solenoid throttling valve  105  functions to control the pressure (and flow) of the low pressure pump to high pressure pump  55  in response to commands from the ECM. This system is functional. However, it has been determined that because of viscosity changes or ranges of viscosity of the hydraulic oil to which the pump is subjected and because of the different flow rates which have to be throttled, solenoid valves of considerable size (having power to infinitely change flow rates over large operating flow conditions at various viscosities) and expense are required. This is so even considering that the solenoid valve is controlling the flow of a low pressure pump and not a high pressure pump. The throttling valve of this invention allows the solenoid valve to be considerably downsized and operate within the broad operating ranges required of a HEUI system. 
     Referring now to FIG. 12, there is schematically depicted throttling valve  200  positioned between low pressure or charge pump  23  and high pressure pump  55  for the HEUI system discussed above. Throttling valve  200  can be viewed as functionally including a flow control valve  202 , a mechanical actuator  203 , a solenoid operated, pressure reducing or control valve  204  and a pressure regulating valve  205 . 
     As discussed, low pressure fluid (at 20 to 60 psi) from charge pump  23  enters inlet  210  of flow control valve  202  at an initial charge pump pressure, P 11 . Flow control valve  202  meters charge pump pressure P 11  to a desired flow control outlet pressure which is outputted at flow control valve outlet  212  and inputted to inlet  106  of high pressure pump  55  at a desired high pressure inlet pump pressure, P 12 . High pressure pump  55  generates high pressure outlet pump pressure P 0  at pump outlet  90  transmitted to the injectors from rail  35 . In the preferred embodiment, for a constant high pressure inlet pump pressure P 12 , high pressure pump  55  produces, at operating pump speeds, a generally constant outlet flow which is at a generally constant high pressure outlet pump pressure P 0 . 
     As schematically indicated in FIG. 12, flow control valve  202  is biased by a spring  213  into, for the preferred embodiment, a full open position. Mechanical actuator  203  opposes the bias of spring  213  and if the mechanical force of mechanical actuator  203  overcomes the bias of spring  213 , flow control valve  202  will be moved into a closed position whereat high pressure pump inlet pressure P 12  will reduce to zero. The force developed by mechanical actuator  203  is a function of the differential in pressure between two fluid pressures exerted at opposite sides or spool ends of mechanical actuator  203 . Fluid at a regulated pressure, P R , is introduced at a closing end  215  of mechanical actuator  203  and the force developed by regulated pressure P R  is counterbalanced by fluid at a control pressure, P C  introduced at a counterbalancing or control end  216  of mechanical actuator  203 . Mechanical actuator  203  controls flow control valve  202  which is thus a slave to the actuator. 
     Regulated pressure P R  is produced at an outlet  218  of pressure regulating valve  205  which is a conventional regulating valve using a preset bias of a spring  219  to drop the pressure of high pressure pump output P 0  introduced to regulating valve inlet  220  to produce regulated pressure P R . Regulating valve  205  does not meter any appreciable flow of fluid from high pressure pump output to drain (not shown in schematic of FIG. 12) and does not materially change high pressure pump output pressure P 0  in rail  35 . If high pressure pump output P 0  drops to an unactuated pressure, i.e., engine shut-off condition, regulating valve spring  219  will open fluid communication between regulating valve inlet and outlet  220 ,  218  so that fluid remains in mechanical actuator  203  at some nominal pressure. 
     Fluid at control pressure P C  is produced at an outlet  223  of pressure control valve  204 . Fluid at regulated pressure P R  from outlet  218  of regulating valve  205  is introduced at an inlet  224  of pressure control valve and metered to a set pressure by a solenoid  225  acting against the bias of a pressure control spring  226 . Solenoid  225  is under control of ECM  18  and has the ability to meter flow through pressure control valve  204  from zero to regulated pressure P R . In event of solenoid failure, fluid communication from regulating valve outlet  218  to control valve outlet  223  is closed thus forcefully biasing actuator  203  and consequently valve  202  to the closed position preventing the supply of oil from pump  55  to rail  35 . 
     In the preferred embodiment and on start-up of a cold engine, high pressure pump output P 0  will be insignificant and fluid connections  220 ,  218  along with fully actuated solenoid  225  and fluid connection  218 ,  223  will place balancing forces on mechanical actuator  203  so that pressure in passages  215  and  216  are equal. Consequently, flow control spring  213  will bias flow control valve  202  into a full open position. Thus maximum flow to high pressure pump inlet  106  will occur. During engine warm-up, high pressure pump  55  will develop sufficient pressure to allow pressure regulating valve  205  to function at which time pressure control valve  204  will likewise function. In the preferred embodiment and in the event of an electrical failure of solenoid  225 , pressure control valve  204  is designed to reduce control pressure P C  to zero with the result that regulated pressure P R  only acts on mechanical actuator  203 . Regulated pressure P R  is set to be sufficient to overcome the bias of flow control spring  213  and close or materially reduce the flow of fluid through flow control valve  202 . The result is then that high pressure pump  55  is starved for fluid and the engine stalls because there is insufficient pressure to operate the fuel injectors. Alternatively, the setting of regulated pressure P R  coupled with the setting for spring bias  213  and the design of flow control valve  202  (as explained below) can be set such that when electrical failure of solenoid  225  occurs, there is sufficient high pressure pump inlet pressure P 12  to allow the fuel injectors to minimally operate. The vehicle could then operate in a “limp home” mode. 
     It should be clear from the discussion of FIG. 12 that there is, for all intents and purposes, an insignificant flow of fluid through pressure control valve  204  and pressure regulating valve  205  or the mechanical actuator  203 . Thus the functioning of the components which regulate flow control valve  202  are isolated from the effects of viscosity or changes in the viscosity of the fluid flowing through flow control valve  202 . Parasitic power losses are also minimized due to minimal flow losses. 
     Further, the regulating pressure P R  (while higher than charge pump pressure P 11 ) is set at a relatively low value when compared to the pump output pressure P 0 . This relatively low pressure lends itself to rapid and responsive modulation through pressure control valve  204 . Solenoid  225  can be selected as a small sized, low cost but truly responsive item. By way of example and not necessarily limitation, in the preferred embodiment, initial charge pump pressure P 11  can range from 0 to 7 bar; high pressure inlet pump pressure P 12  can range from [(0 to 7 bar)−1]; high pressure outlet pump pressure P 0  can range from 0 to 280 bar; regulated pressure P R  is set at a constant pressure established by the relationship of spring  213  and valve  204  (The preferred embodiment utilizes production established components and a 32 bar setting. Other settings are possible.) and the control pressure P C  can vary from 0 to 18 bar. The flow range of low pressure pump is 0-25 Lpm and the viscosity range of the fluid, which in the preferred embodiment is engine oil, is 8-10,000 cSt. 
     Referring now to FIG. 13 there is shown in sectioned view, throttling valve  200  and reference numerals used with respect to discussing the functioning of throttling valve  200  in FIG. 12 will apply to FIG.  13 . Throttling valve  200  shown in FIG. 13 has a first casing section  230  containing flow control valve  202  and a second casing section  231  containing mechanical actuator  203 , pressure control valve  204  and pressure regulator valve  205 . It is contemplated that first casing section  230  may be formed integral with pump housing  56 . Accordingly throttling valve inlet is designated as reference numeral  79  which is the inlet in high pressure pump  55  that is in fluid communication with low pressure pump  23  and throttling valve outlet is designated as reference numeral  106  which is the inlet for high pressure pump  55 . Within first casing section is a drilled passage providing fluid communication between throttling valve inlet and outlet,  79 ,  106 . Within the drilled passage is a cylindrical sleeve  234  and reference may had to FIG. 14 which shows a perspective view of sleeve  234 . In the preferred embodiment, one axial end of sleeve  234  is adjacent throttling valve outlet  106  and the opposite axial end of sleeve  234  is adjacent second casing section  231 . In between the axial ends of sleeve  234  is a plurality of longitudinally spaced orifice openings  235  in fluid communication with throttling valve inlet  79 . The orifice openings permit low pressure pump fluid to flow from throttling inlet  79  through orifice openings  235  into the interior of sleeve  234  and out through throttling outlet  106 . Each orifice opening  235  is dimensionally sized relative to its longitudinal position with respect to throttling inlet  79 . In the preferred embodiment, the largest orifice openings  235  are positioned closest to the closed axial end of sleeve  235 , i.e., adjacent second casing section  231 . 
     Within sleeve  234  is a slidable hollow piston  238  which has a closed end  239  adjacent second casing section  231 . Flow control valve spring  213  has one end seated against hollow piston closed end  239  and the other end seated against throttling valve outlet  106  biasing hollow piston closed end out of sleeve  234  and into contact with abutting second casing section  231 . In this position which is shown in FIG. 13 flow control valve  202  is wide open and maximum flow occurs between throttling valve inlet  79  and outlet  106 . As explained with respect to the discussion of FIG. 12, mechanical actuator  203  under the control of solenoid actuated control valve  204  regulates the position of piston  238  in sleeve  235 . As is well known in HEUI applications, during cold start of the engine, the engine oil has a viscosity significantly different than that when the engine is at normal operating temperature. Further the force to move hollow piston  238  against the flow (i.e., to close) increases as the viscosity increases. It is important to keep the low pressure pump flow at a maximum at the time of cold start and during warm-up of the engine until oil thins to a desired viscosity, even if initial control instructions from the ECM have to be overridden. The sleeve/piston/variable orifice arrangement discussed for flow control valve  202  is somewhat ideal for this application. Specifically, orifice openings  235  can be set to produce a two-staged flow having a first stage which leaves the valve open and sluggish for a limited travel distance and a second stage where the flow can be precisely metered. As the viscosity of the oil thins, the force required to move the valve diminishes and places it into the second stage where it becomes extremely responsive to slight force changes. 
     Those skilled in the art will recognize that many geometrical variations in the sleeve/piston arrangement shown in FIG. 13 are possible. For example, variable orifice openings  235  could be provided in piston  238  instead of sleeve  234 . The positions of throttling valve inlet and outlet  79 ,  106  could be reversed or both could be longitudinally positioned along sleeve  234 . While the variations mentioned are possible and functional, the preferred arrangement for valve stability and valve response is as shown in FIG.  13 . 
     Referring still to FIG. 13, mechanical actuator  203  simply comprises a shuttle or spool  240  sealingly disposed within a drilled passage in second casing  231 . Attached to one end of spool  240  is an actuator plunger  241  in contact with piston closed end  239 . At one end of spool  240  is closing passage  215  which receives fluid at regulated pressure P R  and at the opposite end of spool  240  is control passage  216  receiving fluid at control pressure P C . Pressure in closing passage  215  exerts a force on spool  240  tending to move spool  240  upward in the plane of the drawing shown in FIG. 13 against piston  238 . Pressure in control passage  216  exerts a force on spool  240  tending to move spool  240  downward in the plane of the drawing shown in FIG. 13 out of second casing  231 . Spring bias  213  plus the pressure in control passage  216  acts against the pressure in closing passage  215 . 
     The advantage of a pilot operated (i.e., spool  240 ) valve compared to a solenoid operated flow control valve can now be explained. First as a matter of definition: 
     Q IN =inlet flow from charge pump  23 ; 
     A MV =Area opening of variable orifices  235  in flow control valve  202 ; 
     P R =limited pressure, for example 40 bar, established by regulating valve  205 ; 
     A PV =pilot valve area defined as diameter of spool  240 ; 
     P C =control pressure established by pressure control solenoid valve  204 ; 
     X PV =axial movement of spool  240  (until stopped by spring  213 ); 
     Q PV =flow across variable orifices  235  in sleeve  234 . 
     For throttling valve  200  as defined, the proportionalities producing valve control are as follows: 
     Q IN ˜A MV ; 
     A MV ˜X PV ; 
     X PV ˜ΔP; 
     ΔP=P R −P C    
     For a flow control valve, one must reference the proportionality Q PV ˜ΔP 1/2 . Controlling the flow linearly with respect to current from a solenoid operated flow control valve will then produce a X PV , vs. current curve that is second order. This translates to poor control at the low end of the flow curve in the throttling valve. Utilizing the pilot operated pressure control valve disclosed, one must reference the fact that ΔP=P R −P C . Since P R  is a constant, this relationship is always linear, thus a linear P C  vs. current curve will produce a linear relationship between the current and X PV , this is the preferred control relationship. 
     Pressure regulating valve  205  is conventional and will not be described in detail herein. In FIG. 13, a regulating spool  245  in regulating valve  205  is shown in its free state in which P 0  at regulating valve inlet  220  is less than or equal to P R . As P 0  becomes greater than or equal to P R , the pressure in regulating valve outlet  218  moves regulating spool  245  towards the right as viewed in FIG. 13 against the bias of regulating spring  219 . A land  246  in regulating spool  245  comes in line with a land (not shown) in regulating valve body. As fluid at pressure P 0  continues to leak into regulating valve outlet  218 , regulating spool  245  continues to move towards the right, as viewed in FIG. 13, until a cross hole  247  reaches a position whereat it opens to a spring chamber (i.e., sump). This vents a small amount of oil at P R  from valve outlet  218  moving regulator spool  245  towards the left to its modulated position whereat land  246  aligns with the land in the valve body. 
     Solenoid actuated pressure control valve  204  is also conventional and a conventional solenoid valve is shown in FIG.  15 . The sump drain diagrammatically shown in FIG. 12 is shown as drain port  250  in FIG. 15. A control spool  251  is configured to close or open either control pressure inlet  224  or drain port  250  providing selective communication with control valve outlet  223 . Control spool  251  includes a control spring seat  252  swaged thereto and control spring  226  biases control spool  251  to the right in the plane of FIG.  15 . When current is generated in the solenoid wiring  225  an electrical field moves control spool  251  toward the left in the plane of the drawing shown in FIG. 15 against the bias of control spring  226 . Fluid at regulated pressure P R  enters control inlet  224  and builds pressure in control outlet  223  and also in the “A” direction against control spring  226  to establish flow from control outlet  223  to drain outlet  250  and thereby establish modulation of the control valve  204 . The pressure build in the “A” direction is related to the current level inputted to solenoid  225  and is usually stored in a look-up table in ECM  18  whereby control of pump  55  is effected. 
     An alternative embodiment is illustrated in FIG. 16 which uses similar components as that set forth in the preferred embodiment and the same reference numerals used in describing the preferred embodiment will apply to the alternative embodiment. FIG. 16 is cited as an alternative embodiment only because it discloses a pilot operated throttling valve and in particular a flow control valve regulated by a mechanical actuator as discussed above for FIGS. 12 and 13. In FIG. 16 an orifice  260  is provided between the closing and control ends  215 ,  216  of mechanical actuator  203 . Under static conditions, i.e., when flow control valve  204  is closed (no flow), actuator spool  240  is balanced and flow control spring  213  biases flow control valve  202  into a full open position. However, this alternative embodiment functions during normal operation by solenoid control valve  204  operating to cause a controlled flow of fluid through control end  216  of mechanical actuator  203  through solenoid control valve  204  to drain. The flow of fluid through orifice  260  results in a pressure drop establishing the pressure differential on actuator spool  240  to control the slave flow control valve  202  as described above. The fluid flow through solenoid control valve  204  exposes the solenoid actuated control valve to the viscosity changes of the fluid and the variations in the flow forces which are avoided in the solenoid actuated control valve  204  in the preferred embodiment illustrated in FIGS. 12-15. In the preferred embodiment, solenoid actuated control valve  204  is only controlling pressure, and communication to drain port  250  is only that necessary to establish the desired control pressure P C  so that flow considerations through the valve are insignificant in the “meter in” arrangement of the preferred embodiment. In the alternative “meter out” arrangement flow considerations through solenoid actuated control valve  204  have to be considered in the control valve design and the solenoid sized accordingly. For this reason, the alternative embodiment is not preferred and is simply disclosed to show an alternative pilot valve arrangement which can be used in the inventive throttled inlet pump/throttling valve system applications of the invention. 
     The invention has been described with reference to a preferred and alternative embodiment. Obviously alterations and modifications will occur to those skilled in the art upon reading and understanding the Detailed Description set forth herein. For example, the invention has been described with reference to a HEUI system where it has particular application. To a similar extent, a steering or hydraulic suspension system on a vehicle has similar considerations and a high pressure pump could be installed in such systems. Typically, those systems would not charge the inlet of pump so drain passages (e.g. drain passage  89 ) would not be provided for internal pump leakage. Also, the specifications discuss the throttling valve for use in a HEUI application which place specific demands on the throttling valve that are reflected in the throttling valve design. However, the inventive throttling valve and the inventive throttled inlet pump/throttling valve system disclosed herein can be used in other applications such as power steering pump applications or in unrelated industrial applications. It is intended to include all such modifications and alterations insofar as they come within the scope of the present invention. 
     Various features of the invention are set forth in the following claims.