Abstract:
An electronic engine management unit includes means for actuating each injector individually at a selected different time, and for a prescribed interval, during each cycle of the engine. A high pressure fuel supply pump having a high pressure discharge passage is fluidly connected to the common rail, and to a low pressure feed fuel inlet passage. A control subsystem controls the discharge pressure of the pump between injection events, by diverting the pump discharge so that instead of delivery to the common rail, the flow recirculates through the pump at a lower pressure. This is preferably accomplished by an inlet control passage fluidly connected to the low pressure feed fuel inlet passage, a discharge control passage fluidly connected to the high pressure discharge passage, and a non-return check valve in the high pressure discharge passage, between the discharge control passage and the common rail, which opens toward the common rail. A control valve is fluidly connected to the inlet control passage and to the discharge control passage, and switch means are coordinated with the means for actuating each injector. While the pump discharge passes through the control circuit but immediately before each injector actuation, the hydraulic circuit is substantially closed whereby the pump output pressure rises from the holding pressure to the high pressure. When the pump output pressure reaches the high pressure an injector is actuated.

Description:
This application was filed under 35 USC §371 based on international application No. PCT/US00/04096, which claims benefit under 35 USC §119 (e) of U.S. Pat. No. 60/120,546, filed Feb. 17, 1999. 
    
    
     BACKGROUND OF THE INVENTION 
     The present invention relates to fuel pumps, particularly of the type for supplying fuel at high pressure for injection into an internal combustion engine. 
     Typical gasoline direct injection systems operate at substantially lower pressure level when compared, for example, to IDI or DI diesel fuel injection systems. The amount of energy needed to actuate the high-pressure pump is insignificant in the total energy balance. However, in a system with a constant output pump and variable fuel demands all of the unused pressurized fuel has to be returned into the low-pressure circuit. A good portion of the energy originally used to pressurize the fuel is then converted into thermal energy and has to be dissipated. Even a relatively modest heat rejection (200-500 Watt) will result in fuel temperature increase (especially if the fuel tank is only partially full) and this will further worsen already serious problems resulting from low vapor pressure of a typical gasoline fuel. Because of that a variable output high-pressure supply pump would be very desirable. 
     Furthermore, a speed range of a typical gasoline engine is substantially wider than that of diesel engines (e.g., from 500 RPM at idle to 7000 RPM or higher at rated speed). With variable pumping pressure it would be easier to optimize the injection rate at any engine speed. 
     Several configurations for a direct injection gasoline supply pump are shown and described in U.S. patent application Ser. No. 09/031,859, filed Feb. 27, 1998 for “Supply Pump For Gasoline Common Rail”, the disclosure of which is hereby incorporated by reference. The present invention can be considered as particularly well suited for implementation in one or more of the embodiments shown in said application, as well as variations thereof. 
     SUMMARY OF THE INVENTION 
     According to the present invention, a high pressure pump provides both variable output and pumping pressure modulation. At a first level of control (gross modulation), the pump does not undergo high pressure pumping action, except when needed. At a secondary level of control (micromodulation), at least the frequency of actuation of an electrically operated, (e.g., proportional solenoid), is manifested as pumping pulses which produce the required average high pressure. 
     The invention can broadly be considered as a method for controlling a common rail gasoline fuel injection system having a high pressure supply pump to the common rail, wherein the improvement comprises recycling the pump discharge flow through the pump at a pressure lower than the rail pressure, between injection events, and restoring the discharge flow to the common rail immediately before the next injection event. 
     The invention may be better understood in the context of a gasoline fuel injection system for an internal combustion engine, having a plurality of injectors for delivering fuel to a respective plurality of engine cylinders and a common rail conduit in fluid communication with all the injectors for exposing all the injectors to the same supply of high pressure fuel. An electronic engine management unit includes means for actuating each injector individually at a selected different time, and for a prescribed interval, during each cycle of the engine. A high pressure fuel supply pump having a high pressure discharge passage is fluidly connected to the common rail, and to a low pressure feed fuel inlet passage. A control subsystem controls the discharge pressure of the pump between injection events, by diverting the pump discharge so that instead of delivery to the common rail, the flow recirculates through the pump at a lower pressure. This is preferably accomplished by an inlet control passage fluidly connected to the low pressure feed fuel inlet passage, a discharge control passage fluidly connected to the high pressure discharge passage, and a non-return check valve in the high pressure discharge passage, between the discharge control passage and the common rail, which opens toward the common rail. A control valve is fluidly connected to the inlet control passage and to the discharge control passage, and switch means are coordinated with the means for actuating each injector, for controlling the control valve between a substantially closed position for substantially isolating the inlet control passage from the discharge control passage and a substantially open position for exposing the inlet control passage to the discharge control passage. 
     The invention may also be considered a method for controlling the operation of a high pressure common rail direct gasoline injection system for an internal combustion engine, comprising continuously operating a high pressure fuel pump to receive feed fuel at a low pressure and discharge fuel at a high pressure to a check valve which opens to deliver high pressure fuel to the common rail. Sequentially, each injector is actuated, and after each injector actuation, an hydraulic control circuit is opened upstream of the check valve, whereby the pump discharge passes through the control circuit instead of the check valve, at a decreased pressure from the high pressure to a holding pressure between the high pressure and the feed pressure. While the pump discharge passes through the control circuit but immediately before each injector actuation, the hydraulic circuit is substantially closed whereby the pump output pressure rises from the holding pressure to the high pressure. When the pump output pressure reaches the high pressure an injector is actuated. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     The preferred embodiments of the invention will be described below with reference to the accompanying drawings, in which: 
     FIG. 1 is a schematic of a first embodiment of a gasoline direct injection system according to the invention; 
     FIG. 2 is a schematic of the embodiment of FIG. 1, between injection events; 
     FIG. 3 is a schematic of the embodiment of FIG. 1, during an injection event; 
     FIG. 4 is a diagrammatic representation of the behavior of the rail pressure, pumping pressure, injector command signal, and proportional control valve signal associated with a first control method for the system of FIG. 1, according to the invention; 
     FIG. 5 is a diagrammatic representation of the behavior of the rail pressure, pumping pressure, injector command signal, and proportional control valve signal associated with a second control method for the system of FIG. 1, according to the invention; 
     FIG. 6 is a schematic of a second embodiment of a gasoline direct injection system according to the invention; 
     FIG. 7 is a graphical representation of the theoretical power requirement utilizing the variable delivery and injection pressure of the invention relative to an unregulated pump; 
     FIG. 8 is a schematic of a third embodiment of a gasoline direct injection system according to the invention; 
     FIG. 9 is a diagrammatic representation of the behavior of the rail pressure, pumping pressure, injector command signal, and proportional control valve signal associated with a third control method, for the system of FIG. 8, according to the invention; 
     FIG. 10 is a schematic of another, enhanced embodiment of the system shown in FIG. 8; 
     FIG. 11 is simplified, longitudinal section view of a high pressure pump for implementing the system schematic shown in FIG. 8; and 
     FIG. 12 is a simplified, cross sectional view of the high pressure pump shown in FIG.  11 . 
    
    
     DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     According to the schematic shown in FIG. 1, gasoline is supplied, via feed line  34  and fuel filter  16 , by an electric feed pump  12  at relatively low pressure (under 5 bar, typically 2-4 bar) from the fuel tank  14  to the high-pressure fuel supply pump  18 . From the high-pressure pump  18  gasoline is supplied to the common rail  20  and from the rail  20  to the individual injectors  22   a - 22   d . According to the invention, a control valve  28  in a internal hydraulic circuit  26 , controls the instantaneous discharge pressure of the pump  18 , by diverting and modulating the pressure of the pump discharge flow. 
     In the embodiment of the hydraulic circuit  26  shown in FIG. 1, piston  30  and associated spring  52  provide a bias on ball  50 , thereby blocking flow between pump inlet passage  36 , inlet control passage  40 , and first branch passage  44  on the one hand, and pump discharge passage  38  and discharge control passage  42  on the other hand. An orifice  48  provides fluid communication from the discharge control passage  42  to second branch passage  46 , which is in fluid communication with control chamber  32  within piston  30 . The valve  28 , preferably a proportional control valve, has a valve member  54  having a valve surface which bears against valve seat  55  when the valve is fully closed. With the preferred solenoid type valve operator  56 , the valve member  54  is normally open but closes upon energizing of the solenoid. The timing and duration of solenoid energization, is controlled by the engine management system (e.g., electronic control unit, ECU  58 ), via signal path  60 . Such control includes the distance by which the valve member  54  shifts toward and away from the seat  55  (i.e., the valve stroke), which is adjustable when a proportional control valve is employed. 
     The ECU  58  also controls the solenoids  64   a - 64   d  associated respectively with the injectors  22   a - 22   d , via signal lines  62   a - 62   d . Each injection event is controlled at least as to start and duration. 
     Between the injection events the proportional solenoid valve is substantially open (either completely denergized or at some reduced duty cycle). The pressure in the control chamber  32  will be low and all the fuel displaced by the high pressure pump will be internally recycled through the pump at some reduced pressure level above the feed pressure but below the high pressure for discharge to the rail. In the embodiment of FIG. 1, this holding pressure between injection events will depend mainly on the piston return spring  52  preload and the back pressure in the control chamber. The low pressure of the feed fuel is less than about 5 bar, the high pressure during steady state operation is greater than about 100 bar, and the holding pressure is preferably in the range of about 10-30 bar. These three pressure regions can be discerned in FIG. 2 from the three different line densities in the various flow passages. 
     The substantial closing and substantial opening of the valve increases flow resistance and decreases flow resistance, respectively, of the fuel passing through the control circuit along the valve seat. The flow resistance is controlled by varying at least one of the spacing of the valve member  54  from the valve seat  55  and the frequency of changes in the spacing. When the valve is substantially closed, the space is eliminated so that flow resistance is essentially infinite and no flow passes along the seat. When the valve is substantially closed, a non-zero minimum space is maintained, providing a higher resistance than the rest of the control circuit but permitting a low flow passing along the seat. 
     It should also be appreciated that the piston in the circuit  26  of FIG. 1 is optional, but it acts as a minimum pressure regulator, providing positive torque and “limp home” pressure for the common rail. 
     FIGS. 4 shows the behavior of the rail pressure, supply pump discharge pressure, fuel injector actuation or commend signal, and proportional control valve energizing or commend signal, along a scale corresponding to engine rotation or crank angle  74 , during steady state operation of the system shown in FIG.  1 . Shortly before the desired start of injection (see phase shift  66 ) the duty cycle  68  of the proportional solenoid valve is increased above a base or minimum level  70 , substantially closing the valve member. The pressure in the piston control chamber  32  will increase as more fuel is supplied through the control orifice  48  than the amount of fuel leaving the control chamber  32  along the proportional valve seat  55 . The pressure increase will be gradual because some small amount of fuel is needed to displace the piston and to close or restrict the flow through the proportional valve. Shortly after the desired high pressure level for the rail is reached, any of the injectors, such as  22   b , is switched on and gasoline is delivered into the designated engine cylinder. At the end of the injection event the injector solenoid  64   b  and the proportional valve solenoid  56  are switched off simultaneously and the pumping pressure will be reduced accordingly. 
     FIG. 4 shows the control embodiment wherein the solenoid valve  56  is not fully closed at the end of injection, but is maintained at a low duty cycle to help establish the subsequent holding pressure. FIG. 5 shows another embodiment wherein the solenoid is completely deenergized at the end of the injection event. 
     In both FIGS. 4 and 5 it can be seen that the control valve begins shifting from the substantially open to the substantially closed condition before actuation of an injector, the control valve remains in the substantially closed condition during actuation of that injector, and the control valve returns to and remains in the substantially open condition simultaneously with the deenergizing of that injector. During steady state operation above idle speed of the engine, the injections are discrete events each beginning on a regular time interval, each event having the same duration which is no greater than about one-half the regular time interval. Each injection event has a unique holding pressure interval and control valve actuation event associated therewith, and each injection event has a unique high pressure pumping duration associated therewith. Each control valve actuation event and each high pressure pumping duration has a longer duration than the associated injection event. The injection event, the control valve actuation, and the high pressure pumping duration, all terminate substantially simultaneously. 
     Because the high pressure pump  18  and the rail  20  are separated by a non-return check valve  24  and because there is no demand for fuel between the injection events, the pressure in the rail will remain more or less constant. The rail, however, does not have capacity to store any significant amount of fuel. Even if the desired pressure was reduced in the mean time, the pressure will drop instantly as soon as the injector opens and the injection will take place at a lower pressure level, determined by a reduced pressure in the control chamber of the intensifier piston. The main advantage of the present invention is that there is always some minimum pumping pressure between the injection events, and the pressure prior to the injection increases gradually. As a result, there will be no torque reversals or zero crossings. Therefore, the pump operation will be very smooth and quiet. 
     Although the proportional solenoid valve  28  response is relatively slow, this can be compensated for by selection of proper phase shift  66  and of the actuating frequency of the valve member  54 . Even with a relatively long phase shift there will always be some net energy savings, as is indicated at  72 . Proportional solenoid valves are relatively inexpensive and can be exactly controlled in open mode. 
     As shown in the system  76 FIG. 6, if a faster responding hydraulic circuit  78  is desired, an injector (externally) or an injector-like fast solenoid switching valve (internally)  84  can be used as a substitute for valve  28  of FIG.  1 . Such valve  84  has a hollow body  90  in fluid communication as by annular chamber  94  with one of the inlet control passage  80  or the discharge control passage  82 , a hole  92  in the body, a needle valve member  86  shiftable within the body to open or close the hole as the solenoid  88  operates, and the other of the inlet control passage or the discharge control passage being exposed to the hole. The reduced pressure between the injection events will then depend either from the pressure drop across the switching valve or from a pressure limiting valve which can be installed in series down stream from the switching valve (not shown). 
     FIG. 7 shows an example of power requirements of unregulated versus modulated pump according to the invention. Although theoretical energy saving as shown in FIG. 7 may be diminished because some power is required to operate the solenoid valve, there still will be net positive energy gain. More important, the energy used to operate the solenoid only insignificantly increases gasoline temperature. This is a main objective of this invention, because it allows operation without low pressure fuel return and/or without need for a fuel cooler. If output modulation is required, there will always be energy losses, based on fuel flow and force (pressure) level, regardless of what control system (pressure regulating valve, solenoid spill valve in the rail, mechanism changing the eccentricity etc.) is used. One exception is inlet metering, but this system seems to be too inaccurate, too slow and it generates a lot of acoustic noise. 
     A schematic of the preferred embodiments  96  and  96 ′ are shown in FIGS. 8 and 10, and a schematic of the preferred mode of operation is shown in FIG.  9 . The primed numeric identifiers in FIG. 10 correspond to the unprimed counterparts in FIG.  8  and only the unprimed will be referred to for convenience. FIGS. 11 and 12 show an example of a hardware implementation, in a configuration similar to that described in U.S. patent application Ser. No. 09/031,859. Only the features of the pump  200  necessary to illustrate the present invention are described herein; the disclosure of that application can be referred to if additional details are desired. 
     The pump high pressure output timing is controlled directly by a solenoid valve  104 . During the solenoid off-time the spring  116  biases the valve needle  106  against the hole  112  and associated seat, restricting flow from discharge control passage  102 . This determines the pump pressure between injections. The pressure is preferably maintained at between 10 to 30 bars. This pressure ensures that there are no torque reversals at any given time, and it can also be used for a “limp home” operation of the engine, in case there are problems in the pressure control circuit (faulty pressure transducer, faulty or disconnected pressure control valve etc.). The spring  116  can alternatively be replaced by a spring and ball valve  118  or the like, for biasing the valve member against the valve seat with an equivalent preload, as shown in FIG.  10 . In this embodiment, a bypass passage  120  fluidly connects the pump inlet passage  36  with the common rail  20  downstream of the non-return check valve  24 . Means such as a check valve  122 , are provided in the bypass passage  120  for preventing flow therein except when the pressure in the common rail exceeds a maximum permitted limit. This limits the pressure increase in the rail caused by, e.g., mechanical problems or thermal expansion. 
     The hole  112  of the valve body  110  is exposed to the discharge control passage  102  and the space  114  within the body surrounding the needle member  106  is exposed to the inlet control passage  100 . The pressure control solenoid  108  is energized shortly before any of the fuel injectors are actuated, resulting in a very rapid pumping pressure increase. Injection takes place during this high pressure pumping phase. The spring ( 116 ,  118 ) and solenoid forces then define the instantaneous pumping pressure. The effective flow resistance of the hydraulic circuit  98  and therefor the effect on the discharge pressure of the pump, can be controlled for a given duty cycle (valve member stroke) by controlling the frequency and duration of the strokes. 
     In FIG. 9, the first two valve commands each contain ten equally timed discrete opening and closing strokes over a time interval slightly longer than the respective first two injector commend intervals. The second two valve commands contain six equally timed discrete opening and closing strokes over a time interval slightly longer than the respective second two injector commend intervals. Both the number of closures and the duration of each closure for latter valve commands, are of lesser magnitude than the number of closures and the duration of each closure for latter valve commands. Higher duty cycle means higher pumping pressure and vice versa. The injector commands, the associated pumping discharge pressure to the rail, and the rail pressure can thus be adjusted with considerable flexibility and precision using the preferred control circuit of the present invention. 
     However, the pressure in the rail will remain more or less constant, because at that time there is no demand for fuel and the non-return check valve separates the rail from the pumping circuit. 
     All the fuel displaced by the pump is then re-circulated back into the pump housing at the lower pressure level. The pump remains relatively cool even during extended periods of re-circulation. Because all pumping chambers are always fully filled, pressure increase is almost instantaneous. Despite the constant output variations the pump operation remains very quiet at all speeds. 
     The pump  200  has a housing  202  (which may consist two or more components such as body and cover, etc.). A drive shaft  204  penetrates the housing and carries an eccentric  206  located in a cavity within the housing. A plurality of radially oriented pumping plungers  208  are connected via sliding shoes  212  and actuating ring  214  for radial reciprocation as the eccentric rotates. Feed fuel at low pressure fills the cavity from inlet passage  36  and is delivered via charging passage  216  within each piston to the high pressure pumping chamber  210 . The highly pressurized fuel discharges into passage  38 , where it encounters check valve  24 . The inlet control passage  100 , discharge control passage  102 , injector-type control valve  104 , valve needle member  106 , and solenoid  108  of the hydraulic circuit of FIG. 8 are also evident. 
     In the embodiment of FIG. 10, a split accumulator  124  for the common rail  20  is additionally featured. The selection of the volume of the accumulator is very critical and it is a result of a compromise between two contradictory requirements. A small accumulator volume provides fast response during transients and also fast pressure build up. This is especially important for systems requiring elevated pressure (30 to 40 bar) at cranking, because of low pump output (versus time) and also because generally the leakage tends to increase at low speed. It is, however, far less critical at any of the normal operational points, because of substantial higher speed (ranging from 850+/−RPM at idle to 6000+RPM at rated speed). Large accumulator volume reduces pressure fluctuation (both hydraulic noise and pressure drop during fuel withdrawal). 
     The split accumulator design divides the effective accumulation volume in two portions, separated by two check valves; one no return valve and one valve preset for certain opening pressure, for example 50 bar. The common rail  20  has first and second ends  126 ,  128  and the fuel injectors are connected thereto between the first and second ends. The accumulator  124  has a first end  130  fluidly connected to the first end of the common rail after the non-return check-valve  24  and a second end  132  fluidly connected to the second end  128  of the common rail. A preloaded check valve  134  preset for a particular opening pressure is situated at the first end  130  of the accumulator to receive flow into the accumulator when opened, and is biased in the closed position toward the first end  126  of the common rail. A no return check valve  136  is situated at the second end  132  of the accumulator, to permit flow out of the accumulator and to close toward the accumulator. The preloaded check valve can be set for an opening pressure above 30 bar, only by spring  138  or as a variable dependent on the pressure in passage  140 , which is in fluid communication with the inlet control passage  100 ′. The preloaded check valve is preferably set for an opening pressure of about 50 bar. A pressure transducer  142  may be connected at the second end  128  of the common rail. 
     During cranking the engine is driven by the starter motor at, for example, 100 to 200 RPM. Because of substantial amount of fuel used for injection, the pressure will remain below the opening pressure of the valve  134  and all the fuel supplied by the high pressure pump  18  can be injected. This will lead to rapid engine firing and subsequent rapid speed increase. The engine speed will quickly reach at least idle speed (700 to 900 RPM) and this speed can be sustained by injecting only a fraction of the fuel delivered by the pump. The excess fuel will cause the pressure to increase and ultimately the valve  134  will open and because of active area increase (the back side of the valve is vented into the low pressure circuit via passage  140 ) it will stay open until the engine is shut off again. From that point on, a larger accumulator volume will be available, resulting in reduced pressure fluctuation. During the fuel withdrawal the fuel will be supplied to the smaller portion of the rail  20  from both sides (one portion coming from the pump  18  and the balance coming from the accumulator through the no return check valve  136  (flowing in the reversed direction) providing more uniform pressure signature in the rail.