DE69417846T2 - COMPACT FUEL SYSTEM HIGH PERFORMANCE WITH STORAGE - Google Patents

Abstract

A unitized fuel supply assembly is diclosed including an in-line reciprocating cam driven pump (14) for supplying fuel to an accumulator (12) from which fuel is directed to a plurality of engine cylinders by means of a distributor (16) mounted on the unitized assembly. Dual pump control valves (20) provide fail safe electronic control over the effective pump displacement. One or more injection control valves mounted on the distributor are provided to control injection timing and quantity. The accumulator (12) contains a labyrint of interconnected chambers which are shaped and positioned to produce a minimum overall package size while providing for easy manufacture. <IMAGE>

Description

This invention relates to a fuel system for an internal combustion engine and in particular to an electronically controllable high-pressure fuel pump assembly.

For more than 75 years, the internal combustion engine has been mankind's primary source of motive power. It would be difficult to overstate its importance or the engineering effort that has gone into perfecting it. The art of internal combustion engine design is so mature and well understood that most so-called "new" engine designs are simply designs that consist of a selection of many known alternatives. For example, an improved output torque curve can be achieved simply by sacrificing engine fuel economy. Emissions reduction or improved reliability can also be achieved at an increase in cost. Other objectives, such as increased power and smaller size and/or weight, can be met, but usually at the expense of both fuel efficiency and low cost.

The fuel system of an engine is the component that often has the greatest impact on performance and cost. As a result, a significant portion of all the engineering effort that has been invested in the development of the internal combustion engine to date has been devoted to fuel systems for internal combustion engines. As a result, today's engine designer has an extraordinary range of choices and possible modifications to known fuel system concepts. The design work usually involves extremely complex and subtle trade-offs between cost, size, reliability, performance, ease of manufacture and subsequent compatibility with existing engine designs.

The challenge for current designers has been significantly increased by the need to meet government-mandated emission reduction standards while maintaining or improving fuel efficiency. Given the maturity of fuel system designs, it is extremely difficult to achieve both improved engine performance and reduced emissions through further innovations in fuel system technology. However, with a series of more stringent emissions standards being mandated by the United States Government for the future, the need for such innovations has never been greater. Achieving these standards, particularly those for compression ignition engines, will require significant innovations in fuel systems unless manufacturers are willing to agree to significantly more costly fuel systems and/or engine redesigns. For example, Cummins Engine Company, Inc., the assignee of the present application, currently manufactures two mid-range compression ignition engines designated the B-Series and C-Series (5.9 and 8.3 liter displacements, respectively). These engines use a state-of-the-art pump-line-nozzle (PLN) type fuel system that Cummins purchases from other manufacturers. However, with this type of fuel system, the B and C Series engines will not be able to meet future emission reduction standards imposed by the United States government.

In the totality of known fuel systems, there are several concepts that at first glance appear to be a possible solution to the requirement for greater emission reduction and satisfactory engine performance. However, these systems are inadequate for various reasons, which are set out below.

One possibility is disclosed in US Patent No. 5,042,445. This patent discloses a cam-driven injector unit designed to provide very high injection pressures (30,000 psi (about 206.8 MPa) or more) even at low engine speeds. Such high injection pressures promote better fuel vaporization during injection, which helps to ensure complete combustion and thus lower emissions in the engine exhaust. The implementation of this concept requires an injector unit (defined as a single device combining a fuel injector with a high pressure pump) adjacent to each engine cylinder, with the injector designed to achieve the desired high injection pressure at low engine speeds. The above injector is provided with a hydraulic chamber of variable length to control the timing of each injection event depending on engine conditions. Excessive pressures are prevented in this type of injector at elevated engine speeds by the provision of a pressure relief valve which releases timing fluid during the injection stroke of the fuel injector unit.

Other types of fuel injector units are known that are capable of adequate high pressure injection and sufficiently precise injection to meet some of the performance requirements discussed previously. One example is disclosed in SAE Document No. 911819, which relates to a PDE injector unit developed by Bosch. Another is disclosed in U.S. Patent No. 4,531,672 to Smith, assigned to the assignee of this application.

While the injector units described above are often capable of meeting desired performance requirements, there are significant cost premiums associated with adapting such injectors to existing engine designs. In particular, retrofitting an existing engine, such as the Cummins B-Series or C-Series engines, with one of the injector unit designs described above would require a major overhaul of the engine. Most importantly, if these types of injectors were considered for the B and C engines, it would be found that a redesign of the block, head, front end and all associated parts would be essential. In short, an essentially new engine would be required with an accompanying retooling investment in excess of several hundred million dollars.

Another way of achieving the desired high pressure injection and variable timing required to meet the more stringent emission control standards is disclosed in a fuel system offered by Bosch under the designation PLD. This design option is characterized by the provision of a separate high pressure pump unit associated with each engine cylinder and connected by a short line to a nozzle arranged to inject fuel into the associated cylinder. Each pump unit is individually and separately controlled from the associated nozzle and from all other pump units associated with the engine. The pump units are mounted on the engine as close as possible to the associated engine cylinder for actuation by the engine camshaft. Although this approach offers advantages in fuel system cost and performance resulting from the use of existing engine components and the minimal impact on the head design, significant changes to the engine block would be required. In particular, the block would have to be completely redesigned to accommodate the mounting of the individual pump units along the engine camshaft. Implementing this solution on the B and C engines would involve an investment estimated at around several tens of millions of dollars.

One approach to achieving high performance that requires less redesign is disclosed in U.S. Patent No. 5,096,121 to Grinsteiner. This type of injector unit includes a fluid pressure boosting piston that acts to multiply the pressure of a driving fluid, such as pressurized lubricating oil, by the ratio of the effective cross-sectional areas of the boosting piston, which contacts the driving fluid on its larger, low-pressure side and the engine fuel on its high-pressure side. Such a design provides the ability to achieve many of the desired performance requirements, but substantial redesign of the base engine is still required. For example, the system requires an entirely new cylinder head to accommodate not only the injector, but also the oil reservoir that provides the boost. A separate lubrication circuit or a completely new lubrication circuit must be provided to direct the driving fluid through a control valve to the boost piston. Such a system would require its own intake manifold, oil pump and filtration system.

The cost of redesigning the base engine required by a fluid boost injector unit is likely to be significantly less than the cost of redesigning the engine to accommodate any of the other injector unit and pump unit concepts described previously. Nevertheless, Cummins estimates that adapting B- and C-series engines to fluid boosters would still require an investment in the tens of millions of dollars. In addition to the costs associated with redesigning the engine, the fuel system itself, with the hydraulic injector units, redesigned lubrication circuit, filters, and associated equipment, would likely be far more expensive than many other known types of fuel systems. U.S. Patent Reissue No. 33,270 to Beck et al. discloses another type of hydraulic boost injector unit that appears to provide the same benefits but has the same deficiencies discussed previously.

Another way to achieve the goal of increased fuel system performance would be to provide an accumulator for storing the output of a high pressure pump and to provide a plurality of injectors connected to the accumulator and associated with the engine cylinders, each nozzle containing its own integrated solenoid valve for controlling the timing and amount of fuel flow from the accumulator to each cylinder. Examples of this type of system are disclosed in U.S. Patent No. 5,094,216 to Miyaki et al. and SAE Article No. 910252 entitled Development of New Electronically Controlled Fuel Injection System ECD-U2 for Diesel Engines by Miyaki et al. This system allows the accumulator pressure (and hence the injection pressure) to be regulated on demand independent of engine speed. However, solenoids capable of handling the very high pressure and the necessary short response times are relatively bulky and expensive. Such solenoids require a significant redesign of the head on the C-series and some modification on the B-series engines. Likewise, attaching the high-pressure accumulator to an internal combustion engine is not necessarily straightforward, nor does it result in an orderly engine design or appearance. Although the overall cost of redesigning the engine is less than the engine redesign costs associated with adapting to the above-mentioned fuel systems, the costs associated with the fuel system components themselves, including the high-pressure pump and the electromagnetically controlled injectors, could be prohibitively high.

The approaches described above could potentially meet many of the desired performance requirements, but each design involves a significant cost premium in the form of an expensive engine redesign or additional fuel system costs, or both. Other, less expensive fuel system concepts are known, but these concepts do not meet the desired performance requirements in their entirety.

One solution, which in principle does not require engine redesign, involves the provision of a high pressure in-line pump, such as that offered by Bosch under the designation P7100. In this type of system, injectors located on each engine cylinder are connected by separate lines to corresponding pump chambers contained in the housing of a single high pressure pump. The chambers are aligned along the axis of a pump drive shaft and contain corresponding plungers mounted for reciprocating movement through the pump drive shaft in synchronism with the engine crankshaft. With appropriate design and control, in-line systems of this type can achieve the necessary pressures and injection accuracy under certain engine conditions, but are not reliable in meeting the desired performance requirements over the long term, particularly at low engine speeds. Furthermore, in-line fuel pumps capable of even approaching some of the more important pressure and control objectives are significantly more expensive than the current pump-line-nozzle system used on the Cummins B- and C-Series engines.

Another fuel system that requires a minor redesign of the base engine involves the use of a rotary pump design. This type of pump is characterized by a pump housing that includes a plurality of radially aligned pump chambers in which are installed plungers that can be reciprocated by a cam surface located in the center of the pump housing. U.S. Patent Nos. 4,498,442 and 4,798,189 disclose examples of this type of pump. Although the impact on the engine is small and the cost relatively tively low, the rotary pump lacks performance at higher engine power ratings. In particular, rotary pumps are not able to deliver the desired volume or high pressure across the entire operating range of a typical engine.

Another fuel system concept is disclosed in Japanese Patent Application 57-68532 by Nakao, assigned to Komatsu. In this reference there are disclosed an electronically controlled high pressure pump and an accumulator for receiving the pump output for supplying a plurality of injectors through a manifold-type valve and corresponding fuel supply lines. The timing and amount of injection is regulated by rotary valve elements combined with the manifold valve. The pressure in the accumulator is regulated by a feedback signal dependent on the accumulator pressure to control the effective displacement of the high pressure pump. While this design has interesting features, it does not disclose how the necessary operating pressures can be achieved in a molded assembly of sufficiently compact size so that the resulting system can be conveniently attached to an internal combustion engine. No provision is made to ensure that the system is fail-safe should one or more of the electronic control mechanisms fail during operation. Furthermore, the design provides for a completely separate pump assembly and storage components, which are connected by a number of separate fluid lines, thereby multiplying the potential leak points.

The Komatsu reference also does not teach how to practically manufacture an accumulator so that very high pressures, i.e., from about 34.47 to 206.8 MPa (5000 to 30,000 psi) or more, can be stored in a compact structure with adequate fluid storage capacity without possible leakage or dangerous failure. The Komatsu reference further does not suggest how to design and assemble the system to achieve an acceptably low manufacturing cost. The disclosed distributor valve would also not be suitable for handling the very high pressures required for the system without simultaneously involving a high probability of fuel leakage. causing high parasitic losses, meaning that too much mechanical energy would be required to drive the fuel system pump, which would otherwise be available as useful output from the engine.

Other references have disclosed the concept of providing an accumulator in a fuel system, whereby fuel from the accumulator can be alternatively controlled for injection into the respective engine cylinders either by a distributor valve or a plurality of electromagnets associated with each of the individual injector nozzles. DE 36 18 447 A1, assigned to Bosch, discloses an example of this type of system. The very schematic disclosure of this teaching, however, means that this reference does not indicate how to solve the problems mentioned with reference to the Komatsu reference.

Attempts have been made to construct a high pressure manifold or reservoir for storing the output of a high pressure pump for delivery to the injectors. For example, U.S. Patent No. 5,109,822 to Martin discloses a high pressure manifold fuel injection system having a manifold formed from a one-piece metal housing with a series of elongated bores formed therein to temporarily store the high pressure fuel delivered by a high pressure pump. However, Martin does not teach how to determine the optimum arrangement of the elongated chambers or bores so as to produce a compact manifold with minimal external dimensions that will fit within existing, available mounting panels required by existing engines while ensuring that the manifold housing walls are sufficiently strong to withstand the forces generated by the very high operating pressure of the fuel in the chambers. In addition, Martin does not disclose how to determine the required minimum fuel storage volume for the manifold, which is a primary factor in the design of a compact manifold. Likewise, the manifold disclosed in Martin is not integrated with the high pressure pump unit and/or other components, such as a fuel pump control valve, to provide a compact fuel delivery arrangement. capable of effectively controlling the pressure in the manifold. U.S. Patent No. 2,446,497 to Thomas discloses a high pressure pump, a common high pressure chamber or accumulator, a distributor and fuel injection control regulators mounted side by side to form a combined fuel injection assembly. However, Thomas does not disclose a fuel assembly that is highly compact and integrated and also capable of efficiently and effectively controlling both the pressure in the accumulator and the injection timing and amount.

Attempts have also been made to design high-pressure, high-velocity, electromagnetically operated valves for use in compression ignition fuel systems for internal combustion engines. For example, U.S. Patent No. 3,680,782 to Monpetit et al. discloses an electronically controlled fuel injector that uses a force-balanced three-way valve with a nearly force-balanced "pin-in-sleeve" valve element design. In valves of this type, the movable valve element is movable between first and second positions to alternately connect an output valve passage to one or two alternative valve passages, usually to a high pressure source and a discharge. The movable valve element includes a cavity that opens at one end for telescopically receiving a floating pin. A first valve seat is formed between the sleeve and the surrounding valve housing, and a second valve seat is formed between the sleeve and the pin. The valve element is movable between a first position in which the injector nozzle is connected to a fuel source under high injection pressure and a second position in which the valve element isolates the fuel source from the injection openings of the nozzle and connects the passage leading to the injection openings to an outlet, so that an almost instantaneous termination of each injection event is guaranteed.

Other examples of three-way high-speed, high-pressure fuel system valves are disclosed in U.S. Patent No. 5,038,826 to Kabai et al. (Nippondenso). The "pin-in-sleeve" arrangements of the Monpetit et al. and Nippondenso references, while capable of handling high pressure and high speed operation, allow do not require that the effective valve seats of each disclosed design be substantially unequal in size while the valve element remains substantially force balanced.

Another important feature of an effective fuel delivery system is the ability to regulate injection pressure as needed, independent of engine speed. U.S. Patent No. 5,094,216 to Miyaki et al. and U.S. Patent No. 4,502,445 to Roca-Nierga et al. both disclose a multiple chamber in-line fuel pump assembly with an output control device that varies the effective displacement of one or more pump plungers by providing a separate pump control valve for each pump chamber that is actuated to vary the start of injection, with a constant end of injection occurring when the pump plunger reaches its top dead center position. Specifically, fuel is supplied to the pumping chamber during the return stroke and then pumped out of the pumping chamber during the forward or pumping stroke until the control valve closes and blocks the discharge of fuel from the chamber, thereby starting injection or delivery. The discharge or drain from the pumping chamber is not completed until the end of the plunger's pumping stroke.

Another important feature of an effective fuel delivery system capable of meeting the ever increasing demands for emission reduction is the ability to control the amount of fuel delivered during each injection event. It has been found that the level of emissions produced by the diesel fuel combustion process can be reduced by reducing the volume of fuel injected during the initial phase of the injection event. One method of reducing the initial volume of fuel injected in each injection event is to reduce the pressure of the fuel delivered to the nozzle assemblies during the initial phase of injection. Various devices have been developed for controlling or regulating the amount of fuel delivered during the initial phase of fuel injection in order to reduce the pressure of fuel delivered to the nozzle assemblies. For example, US Patent Nos. 3,718,283, 3,747,857, 4,811,715 and 5,029,568 disclose devices associated with each injector nozzle assembly to provide an initial period of restricted fuel flow and a subsequent period of substantially unrestricted fuel flow through the nozzle orifice into the combustion chamber. However, these flow control devices require changes to each of the fuel injector assemblies in a multiple injector system, thereby making the injection system more expensive and complex. U.S. Patent No. 4,469,068 to Kuroyanagi et al. discloses a distributor-type fuel injector having a variable volume accumulator to vary the amount of fuel injected and achieve effective combustion. However, this device uses a complex accumulator control system to vary the amount of fuel injected, particularly designed for use with a distributor having a reciprocating plunger.

Manifold type fuel injection systems are also subject to other undesirable phenomena known as secondary injection. As the nozzle element of the nozzle assembly closes at the end of each injection event, reverse pressure waves or pulses are generated which travel upstream in the fuel supply lines back to the manifold or dispensing valves. Under certain operating conditions, these pressure waves can be reflected back to the nozzle assembly by the manifold or dispensing valve and generate a secondary nozzle operating pulse of sufficient magnitude to unseat the nozzle valve and cause undesirable secondary injection. U.S. Patent No. 4,246,876 to Bouwkamp et al. discloses a conventional damper valve used to dampen or dissipate the pressure wave energy traveling from the nozzle valve, thereby preventing secondary injection by minimizing the strength of each reflected pressure wave received. However, this design requires a separate damper valve used in each fuel injection line, thereby increasing the cost of the system. U.S. Patent Nos. 4,336,781, 4,624,231 and 5,012,785 each disclose rotary manifold fuel delivery systems that use a single damper-type valve located in the rotary shaft of the manifold for dampening pressure waves in each injection line.

To achieve accurate and predictable fuel injection rates during each injection event, it is necessary to ensure that the fuel delivery circuit connecting the fuel supply to the nozzle assemblies is constantly filled with fuel. It has been found that vapor pockets or bubbles (so-called voids) in the delivery circuit result in insufficient injection pressure and in variations in both the fuel quantity and the timing of the injection. Vapour pockets or bubbles are particularly likely to form in high pressure lines of fuel systems where such lines are connected to a low pressure outlet. If the fuel delivery circuit, and thus an injection line, is connected to an outlet at the end of the injection event, fuel will exit the nozzle at one end of the injection line while fuel will exit to the outlet at the other end of the circuit, rapidly drawing fuel from intermediate parts of the circuit and the injection line and reducing the pressure in them. This effect can lead to the formation of a vapor pocket or bubble in the fuel delivery circuit and the injection line between the spout and the nozzle. Damper valves, mentioned previously in relation to preventing secondary injections, are also used to prevent excessive cavitation by allowing substantially full flow through an injection line to an injector while limiting the back flow of fuel from the injector, thereby retaining fuel in the fuel delivery lines. For example, Japanese Patent Publication 05-180117 discloses a damper valve located downstream of a delivery valve to prevent cavity erosion. The damper valve includes a spring-loaded valve element having an orifice and a pressure regulating valve located in a side channel. The device appears to regulate the fuel pressure in the fuel injection line between the damper valve and a fuel injection valve to a value below a predetermined maximum.

In short, the current state of the art does not provide a practical, cost-effective fuel system that meets the conflicting demands of emission control and improved engine performance. performance, particularly in situations where an existing engine design is to be retrofitted. Furthermore, there are no fuel system components (such as accumulators, solenoid valves, and injection control valves) that possess all of the necessary characteristics to deliver fuel under extremely high pressure in precise quantities at precise times, as determined by controls responsive to a wide range of engine conditions.

The object of the present invention is to remedy the deficiencies of the prior art and in particular to provide a practical, cost-effective fuel system which meets the conflicting demands for emission control and improved engine performance while requiring minimal modification to existing engine designs.

The above-mentioned object is achieved by a fuel pump arrangement according to claim 1. Preferred embodiments are the subject of the subclaims.

One aspect of the present invention is to provide an electronically controllable high pressure fuel pump assembly having a pump, accumulator and distributor in combination with an electrically actuated pump control valve and an injection control valve attached to the modular or unitary assembly. This arrangement allows a highly integrated fuel system to be designed, built and installed on either an original or existing engine design.

Another aspect of the present invention is to provide a fuel system for a compression ignition type internal combustion engine capable of achieving very high injection pressures, i.e. 5000-30000 psi (about 34.47 to 206.8 MPa) and preferably in the range of 16000-22000 psi (about 110.3 to 151.7 MPa), with precise control of the amount and timing depending on different engine conditions.

Another aspect of the present invention is to provide a high performance, high pressure fuel system designed to be retrofitted to existing compression ignition type engine designs without requiring substantial and expensive engine redesign. In particular, the present invention provides a fuel system having the above-mentioned features while also improving engine performance by minimizing parasitic losses despite increasing fuel pressure to a very high level.

Another aspect of the present invention is to provide a highly integrated fuel system characterized by high pressure injection, minimal impact on existing engine designs, precise control of injection quantity and timing, redundant fail-safe electronic components, and improved engine performance at an overall lower cost relative to competing prior art systems.

Another aspect of the present invention is to provide a fuel pump assembly characterized by the combination of a pump, a distributor and an accumulator, the accumulator comprising a housing containing a fluidly connected labyrinth of accumulator chambers sized and positioned relative to one another to provide an ideal integrated device.

Another aspect of the present invention is to provide an improved fuel system capable of delivering sufficiently high operating injection pressures to achieve a significant reduction in emissions, the system incorporating a modular arrangement of sufficiently compact size so that the resulting system can be conveniently fitted to existing internal combustion engines without creating a cluttered, unsightly engine appearance.

Another aspect of the present invention is to provide a fuel system having the above features, wherein the number of fuel leakage points is reduced by reducing system components and providing fail-safe, surplus low-pressure fuel outlets throughout the system to capture any fuel that may leak through primary sealing areas and return it to the fuel system.

Another aspect of the present invention is to provide a fuel pump assembly having a pump housing having a pump cavity oriented in a radial direction, and having an accumulator attached to the pump housing with an overhang in either the transverse and/or longitudinal direction, and a pump control valve disposed on the overhang portion of the accumulator housing adjacent to the pump housing to provide a particularly compact, integrated fuel pump assembly.

Another aspect of the present invention is to provide a fuel pump assembly having a fuel pump that delivers high pressure fuel, i.e., 5000-30000 psi (about 34.47 to 206.8 MPa), and preferably in the range of 16000-22000 psi (about 110.3 to 151.7 MPa), a pump cavity opening into a head engaging surface, and an accumulator configured to receive the output of the pump and temporarily store the fuel at the high operating pressure for subsequent injection into the internal combustion engine, the accumulator being disposed in contact with the head engaging surface of the fuel pump to form an end wall for the pump cavity.

Another aspect of the present invention is to provide a fuel pump assembly having a pump housing having a radially directed pump cavity and an accumulator mounted adjacent one end of the pump housing having at least one chamber and a side dimension such that the accumulator forms an overhang in either the transverse or longitudinal direction perpendicular to the radially directed cavity, in further combination with an injector for directing high pressure fuel to various engine cylinders in time with engine operation, the manifold being cantilevered to the pump housing in spaced relation to the accumulator overhang.

Another aspect of the present invention is to provide a fuel pump assembly having a pump housing having a pump cavity oriented in a radial direction and an accumulator housing secured to the pump housing at one end of the pump housing to form a cantilevered lateral overhang such that the overhang provides an offset cross-section for the fuel pump assembly to complement the irregular cross-section of the internal combustion engine to which the fuel assembly is to be secured.

Another aspect of the present invention is to provide a fuel pump assembly having a pump housing containing a pump cavity, a drive shaft adapted for mounting in the pump housing, a pump head secured to the housing opposite the drive shaft, and a pump unit retained in the pump head by a bracket causing the pump unit to extend into the pump cavity of the pump housing in a spaced, non-contacting relationship with the pump housing, whereby the pump unit can be relatively easily removed and replaced to provide cost effective overhaul of the pump assembly and/or the ability to change pump units to adjust the effective displacement of the fuel pump assembly.

Another aspect of the present invention is the provision of an accumulator for a fuel pump system, the accumulator being formed by a housing containing a fluidically connected labyrinth of chambers, the housing being formed from a one-piece block.

A more specific aspect of the present invention is the provision of a modular or unitary fuel pump assembly for periodically injecting fuel through a plurality of fuel injection lines into respective engine cylinders of a multi-cylinder internal combustion engine. The assembly includes a pump for pressurized fuel and an accumulator for accumulating and temporarily storing pressurized fuel received from the pump. The accumulator is mounted on the pump housing opposite the drive shaft of the pump, with a plurality of pump cavities disposed between the drive shaft and the accumulator. The fuel pump assembly further includes a fuel rail for establishing periodic fluid communication between the accumulator and each of the engine cylinders through the respective fuel injection lines. The fuel rail is secured to the pump housing adjacent one end of the drive shaft and includes an injection control valve for controlling the timing and amount of fuel injected into each cylinder depending on engine operating conditions. The control valve includes a solenoid actuator secured to the rail housing and is oriented generally or substantially in the same radial direction as the pump cavities relative to the axis of rotation of the drive shaft. This arrangement provides an extremely compact, highly integrated fuel pump assembly which maximizes low cost, reduced size and high performance in a fuel system intended for deployment on new or existing engine designs.

Another aspect of the present invention is to provide a modular, unitary, one-piece fuel pump housing including a plurality of outwardly opening pump cavities, a radially enclosed drive shaft, a pump head engagement surface, and a plurality of cam guide surfaces in respective pump cavities, wherein the cam guide surfaces, the head engagement surface, and the drive shaft mounting surfaces are the only surfaces requiring precise machining to provide proper alignment between the drive shaft and the cooperating fuel pumping elements of the pump.

Another aspect of the present invention is to provide a fuel pump having an accumulator, a distributor that supplies fuel to a plurality of engine cylinders, and a pair of associated pump control valves for controlling the displacement of the pump elements to cause the pump elements to share the load necessary to maintain the desired fuel pressure. A first injection control valve is for controlling a pilot injection portion or quantity of injection for each cylinders, and a second injection control valve, associated with the first injection control valve, is provided for controlling a main injection portion or quantity of the injection for each cylinder. Furthermore, an electronic control means is provided which causes an associated valve to take over should one of the (pump or injection) control valves fail.

Another aspect of the present invention is to provide a pump assembly having a pump housing including a pump plunger that reciprocates along a first pump axis, a drive shaft that rotates about a drive axis perpendicular to the pump axis, and an accumulator having at least one elongated chamber secured to the pump housing, the central axis of the chamber being parallel to the drive shaft axis of the pump. This arrangement allows an ideally compact modular or unitary accumulator-type pump assembly to be formed in a minimal component size while providing for an adequate total volume of high pressure fuel.

Another aspect of the present invention is to provide a fuel pump assembly that includes a pump housing having a plurality of pump chambers and a plurality of electromagnetically actuated pump control valves, equal in number to the pump chambers, for controlling the effective displacement of the associated pump plungers operating in each pump chamber. By this arrangement, a pressure signal representative of the pressure of the fuel in the fuel pump accumulator can be used to control the electromagnetically actuated control valves, thereby adjusting the effective displacement of the plungers so that the fuel pressure in the accumulator becomes equal to a predetermined pressure value.

One aspect of the present invention is the provision of two injection control valves for use in a distributor in combination with a fuel pump system constructed in accordance with the present invention, wherein electronic control is provided so that at least at least a "slow-moving" operation of the engine is possible should one of the injection control valves fail.

Another aspect of the present invention is to provide a distributor having an injection control valve for controlling the timing and amount of fuel injected into each cylinder depending on engine operating conditions, the injection control valve comprising a three-way valve operable when energized to connect an axial supply passage in the distributor rotor to a high pressure fuel reservoir and operable when de-energized to connect the axial supply passage in the distributor rotor to a low pressure outlet.

Another aspect of the present invention is to provide a distributor housing arranged to control the flow of fuel through a fuel supply line from an accumulator to each of a plurality of engine cylinders through a pair of three-way valves located in a supply plane transverse to the axis of rotation of a distributor rotor, the three-way valves being housed in first and second valve cavities located on opposite sides of the distributor rotor and connected by supply and discharge passages. The valve cavities are further connected by a rotor supply bore for supplying high pressure fuel to the distributor rotor. The injector is further characterized by a two-way check valve located in the rotor supply bore for preventing fuel supplied from one valve cavity from flowing into the other valve cavity.

Another aspect of the present invention is to provide a fuel pump assembly having cam driven reciprocating plungers driven by respective cams having at least one lobe for causing an associated pump plunger to undergo a forward stroke and a reverse stroke for each revolution of the camshaft, the total number of lobes being selected to produce a pumping action for each injection event.

Another aspect of the present invention is to provide a replaceable pump unit for each of the respective pump cavities in the pump housing constructed in accordance with the present invention, each pump unit comprising a cylinder containing a pump chamber and a cylinder mount for mounting the pump unit in a recess of the fuel pump assembly reservoir. A check valve is provided to allow one-way fuel flow from the pump chamber into the reservoir. The check valve is connected to a disc disposed at one end of the cylinder to form an end wall of the pump chamber. The disc contains both inlet and outlet passages and the mount is provided to form a gap between the cylinder and the disc to provide a path for returning fuel leakage to a fuel supply passage contained in the reservoir.

Another aspect of the present invention is to provide a high pressure fuel pump assembly having an accumulator for storing fuel prior to distribution to respective cylinders in an internal combustion engine through an injector, the accumulator having a total volume sufficient to prevent a drop in fuel pressure of more than about 5-15 percent, and preferably 5-10 percent, during each injection event, depending on factors such as the compressibility of the fuel, the operating pressure of the fuel, the maximum possible required injection volumes, the timing range and injection duration selected for the engine, the maximum effective displacement of each pump unit, the fuel leakage of the system, the compression of the fuel in the fuel lines, and the loss of fuel to the outlet during valve element movement between the fully open and fully closed positions.

Another aspect of the present invention is the provision of an accumulator for the fuel system constructed in accordance with the present invention, the accumulator including an interconnected labyrinth of chambers, the chambers being elongated and cylindrical and arranged in a substantially parallel relationship. The accumulator chambers are ideally arranged so that they intersect a vertical plane through the storage enclosure in a two-dimensional arrangement.

Another aspect of the present invention is to provide a rotary pump and distributor integrated with a single drive shaft assembly to form a compact fuel system assembly capable of accurately and reliably delivering precise amounts of fuel to an engine while minimizing the size and weight of the assembly.

Another aspect of the present invention is the provision of a high pressure fuel pump assembly having a fuel rail with axially slidable plunger valves in combination with a dedicated injection control valve.

Another object of the present invention is to provide a fuel pump assembly having an ultra-compact pump head and integrated pump chamber that minimizes high pressure fuel leakage while reducing the size and weight of the assembly.

Another aspect of the present invention is to provide a range of pump head/accumulator designs to accommodate pump control valves and check valves in various orientations to minimize undesirable fuel leakage, trapped volume, and size and weight of the assembly.

Another aspect of the present invention is to provide a fuel pump assembly with a transversely oriented pump control valve for reducing the trapped volume in the pump head/accumulator to an absolute minimum.

Another aspect of the present invention is to provide a fuel pump assembly having a pump unit and a transverse pump control valve installed in the cylinder of the pump unit.

Another aspect of the present invention is the provision of various accumulator designs to simplify the formation and manufacture of the accumulator while minimizing the possibility of undesirable fuel leakage from the accumulator chambers.

Another aspect of the present invention is to provide a high pressure fuel system with a separately mounted accumulator, thus allowing the accumulator to be located in potentially more convenient/advantageous locations around the engine, while also reducing the size of the pump head, thereby providing a more compact arrangement that can better accommodate the size constraints of particular engines or vehicle designs.

Another aspect of the present invention is the provision of various rod or gap filter mounting concepts for arranging a rod or gap filter in the disclosed system to prevent damage to the system components by small foreign objects.

Another aspect of the present invention is the provision of a quantity control option for controlling the amount of fuel injected during the initial part or initial phase of the injection process by controlling the pressure increase at the nozzle arrangement.

Another aspect of the present invention is the provision of various cavitation control devices for minimizing the formation of vapor pockets or bubbles within the fuel passages of fuel systems, thereby minimizing cavitation-related anomalies in fuel injection metering and timing.

Another aspect of the present invention is to provide a novel high pressure fuel system with flow control and cavitation control devices capable of maximizing the flow control capability of the system while minimizing cavitation.

Another aspect of the present invention is the provision of a single device which enables quantity regulation and at the same time effectively minimises the formation of voids in the fuel passages of the system.

Another aspect of the present invention is to provide cavitation control devices that are both inexpensive to manufacture and simple and easy to attach to a fuel pump assembly.

Another aspect of the present invention is to provide a cavitation control device that can refill the fuel injection lines to each nozzle assembly after an injection event.

Another aspect of the present invention is to provide a cavitation control device that can regulate the fuel pressure in the fuel supply passages during the run-out process to a value above a predetermined minimum, thereby avoiding excessive cavitation.

Another aspect of the present invention is to provide a cavitation control device that can both regulate the pressure in the fuel supply passages during the run-off event and simultaneously refill the passages between injection events.

Another aspect of the present invention is to provide a high pressure coupling having a plurality of integrally formed feed sections for connection to high pressure fuel lines and an opening for controlling flow through at least one of the feed sections. Another aspect of the present invention is to provide a high pressure coupling for effectively connecting high pressure lines of a fuel system while simultaneously providing a suitable housing for a filter.

Another aspect of the present invention is the provision of a high-pressure coupling which enables a simple and cost-effective design of a quantity control device.

Other individual aspects of the invention are apparent from the following chapters "Summary of the Drawing" and "Detailed Description of the Preferred Embodiments".

Summary of the drawing

Figure 1 is a schematic diagram of a fuel system assembly constructed according to the present invention.

Fig. 1a is a schematic representation of a method of constructing a particular fuel system arrangement in accordance with the present invention.

Fig. 1b-1i are schematic representations of techniques for applying the method of Fig. 1a.

Figure 2 is an exploded perspective view of a fuel system assembly constructed in accordance with the present invention.

Figure 3 is an end elevational view of a fuel system assembly constructed in accordance with the present invention.

Fig. 4 is an end elevational view of the opposite end of the fuel system assembly of Fig. 3.

Fig. 5 is a sectional view of the fuel system of Fig. 2-4.

Fig. 6 is a partial sectional view of the fuel system of Fig. 2-5.

Fig. 7 is a side elevational view of an accumulator used in the fuel system assembly of Figs. 2-6.

Fig. 8 is an elevational view of the bottom of the reservoir of Fig. 7.

Fig. 9 is an end elevational view of the reservoir of Figs. 7 and 8.

Fig. 10a-10l are sectional views of the accumulator of Fig. 7 and 8 taken along lines 10a-10l.

Fig. 11 is a side elevational view of a fuel pump housing used in the fuel system assembly of Figs. 2-6.

Fig. 12 is an elevational view of the top of the fuel pump housing of Fig. 11.

Fig. 13 is a sectional view of the fuel pump housing of Fig. 11 taken along line 13-13.

Fig. 14-16 are sectional views of the fuel pump housing of Fig. 11-13 taken along lines 14-14, 15-15 and 16-16.

Fig. 17a is an end elevational view of a distributor housing used in the fuel system assembly of Figs. 2-6.

Figure 17b is a side elevational view of the fuel system assembly according to the present invention showing an alternative mounting arrangement with the distributor shaft oriented perpendicular to the pump drive shaft.

Fig. 18 is a second end elevational view of the distributor housing of Fig. 17a.

Fig. 19 is a side elevational view of the distributor housing of Figs. 17a and 18.

Fig. 20 is an elevational view of the top of the distributor housing of Figs. 17a-19.

Figures 21 and 22 are sectional views of the manifold body taken along lines 21-21 and 22-22 of Figure 17a.

Fig. 23 is a sectional view of the manifold with the solenoid-operated control valves associated therewith taken along line 23-23 of Fig. 20.

Fig. 24-26 are sectional views of the distributor housing taken along lines 24-24, 25-25 and 26-26 of Fig. 20, 18 and 23, respectively.

Fig. 27 is a partial sectional view of the distributor rotor and the surrounding housing taken along a plane transverse to the axis of rotation of the rotor.

Figure 28 is a sectional view of another embodiment of a fuel system assembly constructed in accordance with the present invention.

Fig. 29 is a sectional view of the manifold used in the fuel system assembly of Fig. 28, taken along line 29-29.

Figure 30 is a sectional view of another embodiment of a fuel system assembly constructed in accordance with the present invention.

Fig. 31 is a sectional view of a pump housing used in the fuel system assembly of Fig. 30, taken along line 31-31.

Fig. 32 is a sectional view of the pump housing and accumulator used in the fuel system assembly of Fig. 30, taken along line 32-32.

Fig. 33 is a partial sectional view of the accumulator used in the fuel system assembly of Fig. 30, taken along line 33-33.

Figure 34a is a sectional view of a low pressure accumulator used in the fuel system assembly of Figure 30, taken along line 34-34.

Fig. 34b is a sectional view of a second embodiment of the low pressure accumulator used in the fuel system assembly of Fig. 30 taken along line 34-34.

Figure 35 is a schematic diagram of a hydromechanical embodiment of the present invention.

Figure 36 is a schematic diagram of another embodiment of a fuel system assembly constructed in accordance with the present invention with a rotary pump.

Figure 37 is a sectional view of another embodiment of the manifold of the present invention using sliding spool type valves.

Fig. 38 is a sectional view of the spool valve manifold of Fig. 37 taken along line 38-38.

Figure 39 is a partial sectional view of an alternative embodiment of the fuel system assembly of the present invention.

Figure 40 is a partial sectional view of another embodiment of the fuel system assembly of the present invention.

Figure 41 is a sectional view of another embodiment of a fuel system assembly constructed in accordance with the present invention.

Fig. 42 is a sectional view of the fuel system assembly of Fig. 41 taken substantially along line 42-42.

Fig. 43 is a partial sectional view of the fuel system assembly of Fig. 42 taken substantially along line 43-43.

Figure 44 is a partial sectional view of another embodiment of an accumulator/pump housing assembly constructed in accordance with the present invention, taken along line 44-44 of Figure 45.

Fig. 45 is a partial sectional view of the accumulator/pump housing assembly of Fig. 44 taken along line 45-45.

Figure 46 is a partial sectional view of another embodiment of a pump head/pump housing assembly used in the fuel system assembly according to the present invention.

Figure 47 is a partial sectional view of another embodiment of an accumulator/pump housing assembly used in the fuel system assembly constructed in accordance with the present invention.

Figure 48 is a partial sectional view of another embodiment of a fuel system assembly constructed in accordance with the present invention, with vertically mounted pump control valves.

Fig. 49 is a sectional view of the fuel system assembly of Fig. 48 taken along line 49-49.

Fig. 50 is a sectional view of the accumulator of the fuel system assembly shown in Fig. 48 taken along line 50-50.

Fig. 51 is a sectional view of the accumulator of the fuel system assembly shown in Fig. 48 taken along line 51-51.

Figure 52 is a partial sectional view of another embodiment of a fuel system assembly constructed in accordance with the present invention, showing an accumulator removed.

Fig. 53a is a partial sectional view of the fuel system assembly of Fig. 52 taken along line 53a-53a.

Figure 53b is a partial sectional view of another embodiment of the fuel system assembly of the present invention.

Figure 54a is a partial sectional view of an inlet tube in which a rod or edge filter is disposed that is connected to the reservoir of the fuel system of the present invention.

Fig. 54b is another embodiment of a filter housing for mounting the filter in the fuel system assembly of the present invention.

Figure 55a is a partial sectional view of another embodiment of the high pressure accumulator used in the fuel system assembly of the present invention having a single end plate.

Figure 55b is a partial sectional view of another embodiment of the high pressure accumulator used in the fuel system of the present invention, showing two end plates.

Fig. 5Sc is a plan view of another embodiment of the high pressure accumulator used in the fuel system of the present invention.

Fig. 56 is a partial sectional view of a quantity regulating device of the present invention.

Fig. 57 is a graph showing pressure rate as a function of time during an injection event using the flow control device of Fig. 56.

Fig. 58 is a schematic diagram of another embodiment of a quantity regulating device of the present invention.

Fig. 59 is a graph showing the injection pressure as a function of time, controlled by the devices of Figs. 58 and 60.

Fig. 60 is a schematic diagram of another embodiment of a quantity regulating device of the present invention.

Fig. 61 is a schematic diagram of another embodiment of a quantity regulating device of the present invention.

Figure 62a is a sectional view of a high pressure coupling of the present invention with a filter.

Fig. 62b is a sectional view of a high pressure coupling of Fig. 62a taken along line 62b-62b.

Figure 63a is a sectional view of the injection control valve, booster pump and manifold used in the fuel system assembly of the present invention showing the cavitation control devices.

Fig. 63b is a partial sectional view of the manifold of the assembly shown in Fig. 63a taken along line 63b-63b.

Figure 64a is a partial sectional view of a cavitation control device of the present invention, designated A in Figure 63a.

Figures 64b-64e are partial sectional views of various embodiments of the cavitation control devices used in the fuel system assembly of the present invention.

Figure 65 is a schematic diagram of a cavitation control device incorporated in the fuel system assembly of the present invention.

Figure 66 is another embodiment of a cavitation control device incorporated in the fuel system assembly of the present invention.

Fig. 67 is another embodiment of a cavitation control device used in the fuel system of the present invention.

Fig. 68 is a partial sectional view of the manifold similar to Fig. 63b showing the application of the cavitation control device of Fig. 67.

Fig. 69 is a schematic diagram showing another embodiment of a cavitation control device of the present invention, used in the fuel system of the present invention.

Description of the preferred embodiments

Referring to Fig. 1, the modular or unitary assembly of the present invention is illustrated and may be generally referred to as a "Cummins Accumulator Pump System" (CAPS). As shown in schematic form and generally designated 10, the invention includes a high pressure accumulator 12 for receiving high pressure fuel for delivery to fuel injectors of an associated engine, a high pressure pump 14 for receiving low pressure fuel from a low pressure supply pump 15 and for delivering high pressure fuel to the accumulator 12, and a fuel rail 16 for providing periodic fluid communication between the accumulator 12 and each injector nozzle 11 associated with a respective engine cylinder (not shown). The assembly also includes at least one pump control valve 18, 19 disposed along the fuel supply line to the pump 14 for controlling the amount of fuel delivered to the accumulator 12 so that a desired fuel pressure is maintained in the accumulator 12. Likewise, one or more injection control valves 20 arranged along the fuel supply line from the accumulator 12 to the manifold 16 are provided for controlling the timing and amount of fuel injected into each engine cylinder depending on the engine operating conditions. An electronic control module (ECU) 13 controls the operation of the pump control valves 18, 19 and the injection control valve 20 based on various engine operating conditions to ex act, thereby effectively controlling fuel timing and metering.

The injection quantity regulation can be changed by a device located between the accumulator and the distributor.

Figures 2-4 show the preferred embodiment of the fuel delivery assembly 10 in its practical form in a modular or unitary, compact arrangement with an accumulator housing 34 of the accumulator 12 and a distributor housing 44 of the distributor 16, both of which are attached to a pump housing 22 associated with the pump 14. As shown in Figures 11-16, the pump housing 22 includes a lower portion 23 defining a drive shaft receiving cavity 24 for radially enclosing a drive or cam shaft 26. The pump housing 22 also includes an upper portion 25 which is integrally formed with the lower portion 23, for example by metal casting techniques. Two substantially cylindrical pump cavities 28 and 30 formed in the upper portion 25 extend radially from the longitudinal axis of the camshaft 26. The pump cavities 28 and 30 have substantially parallel central axes to form an "in-line pumping arrangement." The upper portion 25 of the pump housing 22 includes a partition wall 31 for separating the pump cavities 28 and 30 and a head engagement surface 32 for engaging the accumulator 12 to form an end wall for the pump cavities 28 and 30. Four openings 33 are formed in the upper portion 25 for receiving bolts (not shown) by which the accumulator housing 34 is secured to the pump housing 22.

The accumulator housing 34 is substantially rectangular in both lateral and vertical cross-section and includes a lower surface which is secured to the head engagement surface 32 of the pump housing 22. Referring to Figs. 5-10a, four recesses 35 are formed in the lower surface of the accumulator housing 34 opposite the corresponding openings 33 which contain internal threads for engagement with complementary threads formed on screws (not shown) which extend upwardly from the openings 33 of the pump housing 22 to secure the accumulator housing 34 to the pump housing 22. housing 22. The accumulator housing 34 includes elongated accumulator chambers 36 extending along the longitudinal direction of the housing 34 for receiving and temporarily storing high pressure fuel discharged from the pump 14. The accumulator housing 34 extends axially outward from the pump housing 22 parallel to the longitudinal axis of a camshaft 26 to form a cantilevered axial overhang 38 with respect to the pump housing 22. Preferably, the central axis of each accumulator chamber 36 is substantially parallel to the drive axis of the camshaft 26 and perpendicular to the pump axis extending in the radial direction through the pump cavities 28 and 30. The accumulator housing 34 also extends laterally outward from the pump housing 22 to form a cantilevered lateral overhang 40. A first pump control valve 18 and a second pump control valve 19 are mounted on the cantilevered lateral overhang 40 of the accumulator housing 34 adjacent to the pump housing 22. As shown in Figs. 2, 3 and 6, the pump control valves 18 and 19 are seated in downwardly open recesses formed on the underside of the accumulator housing 34. In addition, a pressure sensor 42 for determining the fuel pressure in the accumulator chambers 36 is mounted in a recess formed on the underside of the cantilevered axial accumulator overhang 38.

2, 3 and 5, the distributor housing 44 of the fuel distributor 16 is cantileveredly mounted to the pump housing 22 adjacent to the drive shaft cavity 24 and extends outwardly from the pump housing 22 in a spaced, substantially parallel relationship to the axial overhang 38 of the accumulator housing 34. A first injection control valve 20 and a second injection control valve 21 are mounted to the distributor housing 44 in the space between the distributor housing and the cantilevered axial overhang 38 of the accumulator housing 34.

As previously described, the various components of the modular or unitary fuel delivery assembly 10 are oriented in a particular arrangement with respect to one another so that the subsequent connection of the respective housings 22, 34 and 44 provides a compact modular or unitary assembly having axial, radial and lateral external dimensions within which other components such as a pressure sensor 42, injection control valves 20 and 21, pump control valves 18 and 19 and various fuel passages can be easily and effectively integrated into the assembly while maintaining the functionality of each component and the compact design of the assembly.

Referring to Figures 7-9 and 10a-10l, the accumulator housing 34 is formed from a one-piece block made of high strength material such as SAE 4340, VIMVAR grade, tempered to 700 F (about 371ºC); SAE 4140, VIMVAR grade, tempered to HRc 37 and gas nitrided; Maraging 18Ni(250) age hardened to 900 F (about 482ºC); Customer 455 stainless steel age hardened to 950 F (about 510ºC); and Aermet-100 age hardened to 900 F (about 482ºC). The storage chambers 36 are formed in the storage housing 34 by drilling axial bores in the unitary block from an end face of the block. The storage chambers 36 are arranged to intersect a vertical plane extending through the storage housing 34 in a two-dimensional array having an upper row 54 (Fig. 9) of four storage chambers 36a, 36b, 36c and 36d and a lower row 56 (Fig. 9) of three storage chambers 36e, 36f and 36g, as shown in Fig. 9. Each storage chamber 36 is elongated and cylindrical in shape and is arranged adjacent to and in a substantially parallel relationship to another chamber. Likewise, the open end of each chamber 36 is fluid-tightly closed with a plug 58 disposed in a recess 60 formed in the open end. The opposite end of each chamber 36 terminates in the block at a point just before the end point of the elongated extent of the housing 34.

Referring again to the details of the accumulator construction, as shown in Figures 7-9 and 10a-10l, the upper row 54 of chambers 36a-d are fluidly connected by a first transverse passage 62 and an axial passage 64. The first transverse passage 62 extends laterally through the housing 34, perpendicular to the central axis of the chambers 36, to connect the chambers 36b-d of the upper row 54. The axial passage 64 extends perpendicular to the first transverse passage 62, axially along the housing 34 into communication with the chamber 36a, which is shorter than the remaining chambers of the upper row 54. The first transverse passage 62 is formed by drilling laterally through one side of the block to intersect the chambers 36b-d of the housing 34. The open end of the first transverse passage 62 is fluid-tightly closed by a plug (not shown) disposed in a recess 68 similar to the plug 58 and recess 60 of the storage chambers 36. The chamber 36a has been shortened to accommodate the recess 68. The axial passage 64 is formed by drilling from the open end of the storage chamber 36a prior to insertion of the plug 58. Similarly, the storage chambers 36e, 36f and 36g of the lower row 56 are connected by a second transverse passage 69 which is drilled laterally through the housing 34 from one side of the housing 34 and terminates at the chamber 36g. A plug (not shown) is screwed into a recess 69a formed in the open end of the second transverse passage 69 to fluid-tightly close the passage 69. The upper row 54 and the lower row 56 are connected by a vertical passage 71 and an axial passage 73. The vertical passage 71 (Fig. 10b) extends upwardly from the lower surface of the cantilevered axial overhang 38 to connect to the storage chamber 36a. The open end of the passage 71 is fluid-tightly closed by a plug (not shown) disposed in a recess formed in the open end. The axial passage 73 communicates at one end with the storage chamber 36g and at the opposite end with the vertical passage 71. In this manner, the first and second transverse passages 62 and 69, respectively, and the axial passages 64 and 73 interconnect the storage chambers 36a-g to form a fluidly connected labyrinth of chambers for temporarily storing fuel discharged from the pump 14. A fuel supply passage 67 extending from the lower surface of the axial overhang 38 communicates with the storage chamber 36d. A recess formed in the open end of the fuel supply passage 67 serves to accommodate a fuel supply pipe for supplying the temporarily stored fuel to the fuel injection control valves 20 and 21.

Referring to Figures 7, 8, 10b and 10d-10f, the accumulator housing 34 also includes a first pump control valve recess 70 and a second pump control valve recess 72 formed in the lower surface of the housing 34 for receiving the first and second pump control valves 18 and 19, respectively. The first and second pump control valves 18 and 19 are each preferably a solenoid-operated valve assembly of the type disclosed in commonly assigned U.S. Patent No. 4,905,960 to Barnhart, which is incorporated herein by reference. A corresponding valve cavity 74, 76 extends upwardly from each pump control valve recess 70, 72, respectively, but terminates below the accumulator chamber 36a for receiving a control valve element 75 (Fig. 6) of the first pump control valve 18. Two fuel supply side channels 78 and 80 are formed laterally inwardly by a bore from the vertical side of the axial overhang 38 adjacent to the first and second pump control valves 18 and 19, respectively. The open ends of the fuel supply side channels 78 and 80 are each fluid-tightly closed with a plug (not shown) disposed in corresponding recesses formed in the open ends. Each fuel supply side passage 78, 80 communicates with a corresponding valve cavity 74, 76 and extends laterally through the housing 34 and terminates at a position above the corresponding pump cavity 28, 30 when the accumulator housing 34 is secured to the pump housing 22. Additionally, the accumulator housing 34 is provided with a stepped recess 79 (Fig. 10i) formed in the lower surface of the axial overhang 38 adjacent to the second pump control valve recess 72 for receiving the pressure sensor 42. A passage 81 connects the recess 79 to the accumulator chamber 36a.

The reservoir 12 also includes a first pump unit recess 82 and a second pump unit recess 84 formed in the lower surface of the housing 34 in alignment with corresponding pump cavities 28 and 30 of the pump housing. The pump recesses 82 and 84 communicate with and are aligned with the pump cavities 28 and 30, respectively, so that the corresponding pump units 86 and 88 can be inserted into the corresponding pump cavities 28 and 30 and Recesses 82 and 84 can be installed as shown in Figs. 5 and 6. In this way, the accumulator housing 34 and the corresponding recesses 82 and 84 form a pump head for closing and sealing the cavities 28 and 30. First and second pump unit outlet passages 83 and 85, respectively, extend vertically through the accumulator housing 34 and connect the first and second pump unit recesses 82 and 84, respectively, to the accumulator chamber 36c.

A common fuel supply passage 90 (Figs. 5, 10b and 10e) extends laterally inwardly from the vertical side of the side overhang 40 between and parallel to the fuel supply side channels 78 and 80. Two connecting passages 92 and 94 connect the fuel supply passage 90 to the pump control valve recesses 70 and 72, respectively. The opposite end of the common fuel supply passage 90 is connected to the pump recesses 82 and 84 via recess drain passages 96 and 98, respectively (Fig. 10e) to drain leakage of fuel from the recesses 82 and 84, as described in more detail below. The innermost end of each fuel supply side passage 78 and 80 is connected to the corresponding pump unit recess 82 and 84 through fuel passages 100 and 102, respectively (Fig. 10f). In this way, fuel entering the common fuel supply passage 90 flows through the connecting passages 92 and 94 and valve recesses 70 and 72 into corresponding fuel supply side passages 78 and 80 for delivery to the pump units 86 and 88 via fuel passages 100 and 102, depending on the position of the corresponding pump control valves 18 and 19.

The storage chambers 36 are particularly dimensioned to provide a sufficient total volume to deliver a controlled amount of fuel at a predetermined operating pressure to each engine cylinder at appropriate times throughout the operating range of the engine, while minimizing the physical dimensions of the storage housing 34 and ensuring that the walls of the storage housing 34 are sufficiently strong to withstand the forces created by the very high operating pressure of the fuel in the storage chambers 36, e.g. 5000 psi to 30000 psi (about 34.47 to 206.8 MPa) and preferably 16000 -22000 psi (about 110.3 to 151.7 MPa). Determining the minimum fuel storage volume required for a designed accumulator is important in applying the present invention to a particular engine. Storage volume is also related to the sizing of other components. For example, the amount of fuel, timing range, injection pressure, and duration required by an engine are primary factors to be considered in determining the proper size of components used in designing a fuel system according to the present invention, which may be referred to as the "Cummins Accumulator Pump System" (CAPS). As an example, the sizing method for designing a fuel system according to the present invention for application to B and C Cummins engines is described below.

The maximum nozzle pressure for this application was chosen to be 21,000 psi (about 144.8 MPa) with a theoretical duration of 30 degrees crank angle. The accumulator size was determined based on the further constraint that the maximum fuel pressure drop during an injection event should not exceed five percent. The pumping element diameter and stroke were determined by calculating the fuel exchange requirements in the accumulator based on the fuel injection plus losses due to valve transition and leakage, manifold leakage, pumping element leakage, and the injection line volume diverted to the run-off at the end of injection. Since each injection event includes one exchange pumping event (the total number of cam lobes equals the number of engine cylinders), all of the fuel lost from the various sources during an injection should be replaced by the one pumping event.

Another limitation was the maximum allowable energy loss due to leakage and other causes, based on the requirement that the parasitic power of CAPS should not exceed that of conventional types of state-of-the-art in-line pump design when operating at the same injection pressures. Other conditions, such as limitation of pump stroke, leakage and valve transition losses, etc., were taken into account, limiting the size of sealing surfaces for the injection control valve and manifold, and valve transition velocities (to prevent excessive accumulator leakage to the outlet). Once the size of the manifold, valve, accumulator volume, and pumping element stroke were determined, adequate information was available to design the cam, bearings, tappet rollers, and pumping element springs. Finally, to determine the final CAPS component design, the combination of these elements was aligned, rearranged, examined for vehicle and engine interference, and analyzed for acceptable operating load values. Fig. 1a schematically summarizes this design process.

With respect to the accumulator, the following information summarizes the analytical procedure used to determine the minimum required volume of accumulator used in a fuel system constructed in accordance with the present invention for the B and C Cummins engines:

Step 1. Calculate to determine the maximum allowable flow for CAPS pumping elements. Note: The energy to support the flow through the CAPS system should not significantly exceed conventional high pressure high efficiency type PLN fuel systems. Current PLN systems operating at 1200 bar (120 MPa) pump pressure require 5.65 kW drive power at 2400 rpm. Thus, the drive power for the CAPS system should not be significantly greater. Since the pump pressure in the CAPS system is approximately constant, the maximum allowable pump flow can be calculated from the following relationship for a six-cylinder engine:

where: Pwr = power requirement (w)

Np = pump speed (rpm)

P = pump discharge pressure (Pa)

V = pump delivery volume (m**3)

With the design constraint that the power requirement of the CAPS system should not exceed 5.65 kW, this equation can be used to calculate the maximum pump flow rate. At 1100 bar (100 MPa) and 2400 rpm, this calculation indicates that the pump flow rate should not exceed 428 mm³/stroke.

Step 2. Calculate to determine that the CAPS parts or components do not exceed the allowable current and drive force requirements.

The pump discharge volume is the sum of the fuel volumes required for combustion, line pressurization and leakage. Reducing leakage is therefore critical to the successful implementation of the present invention. The leakage volumes were analyzed and reduced through design optimization. The following Table 1 lists the volume contributions to the total pump discharge for a Cummins C-Series engine. Table 1 C-Engines Pump Discharge Breakdown in mm³ for CAPS

* Note: see the following method for leakage loss calculation.

This analysis shows that the CAPS system should not exceed PLN systems at maximum torque through rated speeds with the same injection pressure. At lower speeds, the pump delivery rate increases due to the longer time available for leakage. This volume must be used in the design because high pressure capability at low speed is critical to the CAPS concept. The pumping capacity required at low speeds can be assumed to be higher than conventional PLN systems when the CAPS system is operated at high pressure and low speed.

Step 3. Calculate to determine the storage volume necessary to ensure that the storage pressure does not drop by more than 5% between pumping operations.

Determination of storage volume requirement

The calculation of the storage volume required for a given pressure value and pressure drop during pumping would be calculated as follows. A steady state, uniform flow during the pumping process is assumed for a pumping operation as shown in Fig. 1b.

It is also assumed that the pumping element and fuel supply (injected + leaked) do not occur simultaneously (output flow rate is zero), occur adiabatically and no volume work occurs. Therefore, the energy equation is reduced to the following ratio for a control volume with one inlet.

mihi = m&sub2;u&sub2; - m&sub1;u&sub1;

From the conservation of mass

m₂ = m₁ + mi

and the thermodynamic equation

is to be used

For a low pressure drop, a constant density, a negligible energy content of the intake mass compared to the energy stored in the accumulator and a negligible temperature increase due to the intake fuel quantity are assumed.

Therefore

is converted to volume

m = ρV and mi = m₂ - mi = Δm

where:

P = initial pressure

ΔV = pump volume flow rate per stroke

ρ₁ = density at pressure

u₂ - u₁ = internal energy for fuel

The internal energy of diesel fuel is calculated from the ratio of the bulk modulus as a function of pressure.

where:

C = (B₀/B)1/b

B₀ = Bulk modulus at atmospheric pressure

B = compression modulus at actual pressure

P = pressure

a = constant

b = constant

ρ₀0 = density at atmospheric conditions

The final result is: Equation A

For a given volume change, pressure and pressure drop, the required volume can be easily calculated. As the pump flow increases, the storage volume increases and so the highest pump flow must be used to determine the size of the storage. As shown, the highest pump flow occurs at low speed due to leakage. Using the pump flow of 501 mm³ at low speed and a design condition of 5% pressure drop, the required storage volume is calculated to be approximately 130,000 mm³.

As previously stated, the pump delivery per stroke is the sum of the combustion, the line volume pressurization and the leakage fuel quantity.

AV = VInjected + ΔVLine + ΔVSolenoid Leakage + ΔVManifold Leakage + ΔVPump Leakage

The line volume loss was calculated from the particular energy ratio presented previously. Once the compression energy required to raise the total line volume to the injection pressure was known, an effective fuel volume for a constant pressure was calculated as shown in Fig. 1c and Fig. 1d.

The leakage loss for the solenoid, distributor and pumping element was calculated using energy conservation, pressure vessel expansion formula and diesel fuel thermodynamic properties. The interstitial leakage current can be calculated from the following equation.

where:

D = shaft diameter

h = gap or clearance

ΔP = pressure drop

u = viscosity at temperature and pressure

L = Seal length

Since the temperature profile, viscosity, pressure profile and clearance are unknown and interdependent, the flow is solved iteratively at dx intervals along the seal length, assuming that the enthalpy is constant. See Fig. 1e.

The solenoid valve is more complex due to the parallel current that has to be iterated. Likewise, the valve dynamics are calculated using a spring, mass and damper system with multiple degrees of freedom.

Once the pump displacement was known, the pumping element stroke was calculated using the known plunger diameter. The choice of plunger diameter and stroke involved several iterations of hydraulic force, contact stress, bearing load, instantaneous torque, cam diameter, roller diameter and non-tracking (component inertia). All of these parameters are dependent on the plunger diameter and stroke combination. Optimizing one parameter is likely to adversely affect other parameters. A spreadsheet program can be used to analyze the various design options.

Determination of storage size and shape for a storage volume of 130000 mm³ (Part I)

The CAPS size is determined by enclosure conditions for engine and vehicle components. It was assumed that the same gear train system in the current engine is suitable for driving the CAPS fuel pump. The camshaft, which transmits the power from the gear train to the CAPS fuel pump, was specified as one of the constraints for positioning the CAPS assembly. Fig. 1f shows the constraints for the CAPS assembly as applied to a Cummins engine.

In Fig. 1f, the right and lower surfaces are limited by the engine block. The engine size and other vehicle components limit the left and upper surfaces. (These two surfaces are drawn based on the gear train housing boundary in Fig. 1f). The limitation of the shroud length is determined by the distance between the gear train housing and the engine fuel filter.

Fig. 1g shows how the CAPS assembly fits into the confined enclosure. To avoid contact with the engine block at the top corner, the entire assembly is rotated 30º when installed in the engine. Both side boundaries and the top boundary are tight in the CAPS design planned for the Cummins C-Series engine. However, space is available longitudinally and downwards.

The design shown in Fig. 1g and Fig. 1h was determined by examining numerous accumulator designs. The accumulator dimensions required for a sufficiently strong accumulator consisting of a single internal chamber were determined. It was found that the length of the accumulator did not meet the enclosure requirements. The next step involved examining multiple chamber designs, with some designs including stacked chambers. The multiple chambers increased the width and shortened the length. The addition of stacked chambers decreased the width with some increase in height. The combination of strength, width and length requirements was best met by the multiple stacked chamber accumulator shown in Fig. 1h. The dimensions identified in Fig. 1h are listed in Table 2 below.

Table 2

Dimension Size (mm ± 0.05)

a212

b106

c54

d41

e15

f15

g41

h67

i93

The layout plan of the cylindrical holes was based on (1) the amount of fuel (130000 mm3) contained inside the accumulator as calculated using equation A and (2) the avoidance of fatigue fracture during testing and operational trials. Two rows of cylindrical holes are provided to avoid the long and large holes. Hole No. 1 is shorter than holes No. 2, 3 and 4 to provide sufficient wall thickness to the 4 mm cross-bore plug seat. The lower holes are not designed for the conditions in shorter in relation to the pressure sensor and the fuel pump inlet. All holes are designed for 13 mm diameter and are connected by a 4 mm cross hole or a vertical side hole. The hole dimensions, as shown in Table 3 below, are dimensioned to obtain the desired fuel volume in the accumulator. Table 3 Accumulator hole size

Approximate total storage weight (lbs): 18.82 (18.82 lbs is approximately 8.537 kg)

The wall thickness around the bores is determined so that the stresses at stress concentrations are less than the allowable material strength to prevent fatigue failure. The pressure vessel formula as well as detailed finite element analysis are used to estimate the stress values. Since stress concentration at bore intersections is a key consideration in accumulator design, detailed finite element analysis would provide appropriate local stress results. It is known that the stress concentration factor for closed-end, side-bored or cross-bored cylinders is usually 3.0 to 4.0. For example, the stress concentration factor in Peterson's book is 3.42 for the bore size given in Table 4.

The analytical pressure vessel formula for the maximum tensile stress at in the circumferential direction is

?t = P (b² + a²) / (b² - a²) (1)

where p is the internal radial pressure, a is the cylinder inner radius and b is the cylinder outer radius. The cylinder wall thickness t is calculated by t = b-a. Note that equation (1) applies to cylindrical thick vessels without cutting holes. Likewise, the effect of a closed end cap is not taken into account.

The task is to determine the minimum wall thickness for a specific operating pressure and bore diameter and for specific material properties. Five materials were considered for the production of a storage prototype. These were:

1. SAE4340, VIMAR grade, tempered at 700 F (about 371ºC).

2. SAE 4140, VIMVAR grade, tempered to HRc 37 and gas nitrided.

3. Maraging 18Ni(250), age hardened at 900 F (about 482ºC).

4. Customer 455 stainless steel, hardened to 950 F (about 510ºC).

5. Aermet-100, cured at 900 F (about 482ºC).

Table 4 below shows the wall thickness requirement for various materials and stress amplification factors (SIF) at the bore interfaces. In Table 4, the allowable material tensile stress is calculated from the Goodman diagram for R = O. The stress amplification factor at the bore intersection depends on the bore diameter, intersection angle, bore offset, radius at intersection edge, etc., and the SIF is presented in Table 4 in the form of design input data. The allowable maximum tensile stress within the pressure vessel is the allowable material tensile stress divided by the stress amplification factor. The accumulator drawing shown in Fig. 4B has a minimum wall thickness of 6.5 mm. From the results calculated in Table 4, it is concluded that the wall thickness around the bores is adequate for the chosen material in the accumulator design (1 ksi is approximately 6.896 MPa). Table 4 Sizing of the storage wall thickness

Note: ¹ Working pressure 1350 bar = 19,575 ksi.

² A 0.72 surface finishing factor is included in the fatigue strength

When investigating stresses at the accumulator well intersection, the following types of loads were considered (1 bar corresponds to 100 kPa).

Condition 1: A significant number of engine start/shutdown cycles occur during the accumulator lifetime. This results in an estimated 25000 pressure cycles in the accumulator from 0 to 1100 bar.

Condition 2: Small pressure fluctuations occur during operation in the storage cylinders. A maximum pressure drop of 15% from the maximum pressure value (1100 bar) is assumed. It is assumed that these pressure fluctuations occur from 935 to 1100 bar with a frequency of 108-109 cycles.

A 3-D finite element model is shown in Fig. 1i. The model has 1168 elements and 1566 nodes. The analysis results are summarized in Table 5. The stress amplification factor in the range of 3.0 to 4.4 is estimated for different bore sizes. The Aermet-100 material properties are used to calculate the fatigue span in Table 5. The analysis results in Table 5 show that the accumulator has excellent structural durability when the operating pressure condition does not exceed 1100 bar. Also, air jet flow machining is recommended to improve the intersection geometry and keep the stress concentrations to a minimum, thus preventing fatigue fractures. Table 5 Stress analysis results of accumulator bore intersections

Note: * 1100 bar = 15.95 ksi

* * The material Aermet-100 is used to estimate the fatigue range.

Reference is now made to the details of the pump assembly. In particular, pump units 86 and 88 will now be described in detail with reference to Figs. 5 and 6. Pump units 86 and 88 of pump 14 are of the same construction and therefore only pump unit 86 will be described below. Pump unit 86 includes a pump mount 104 disposed in pump unit recess 82. and extends outwardly toward the camshaft cavity 24. The pump mount 104 has a generally cylindrical shape to form a cavity 105 and includes an upper portion 106 having external threads for engagement with complementary threads formed on the inner surface of the pump unit recess 82. The mount 104 also includes a smaller diameter lower portion 108 that extends into the pump cavity 28 and terminates to form a bottom wall 110. The pump unit 86 also includes a disk 112 disposed within the cavity 105 and the pump unit recess 82 and a pump cylinder 116 mounted adjacent the disk 112 within the cavity 105 of the mount 104. The bracket 104 holds the cylinder 116 and the disk 112 in a compressive abutting relationship, with the disk 112 being pressed against the accumulator housing 34 when the bracket 104 is fully threaded into the recess 82. A central bore 118 extending through the entire length of the pump cylinder 116 is aligned with a central opening 120 in the bottom wall 110 of the bracket 104. A pump plunger 122 is inserted for reciprocating movement in the central bore 118 and central opening 120 to form a pumping chamber 124 between the upper end of the plunger 122 and the disk 112 which forms an end wall 114 for the pumping chamber 124. Thus, the bracket 104 enables the pump unit 86 to be secured within the pump unit recess 82 of the accumulator housing 34 so that it extends into the pump cavity 28 of the pump housing 22 without being in direct contact with the pump housing 22. This arrangement limits the high pressure sealing surfaces to the contact surfaces between the disc 112 and the recess 82 and between the disc 112 and the cylinder 116, thereby eliminating the need for sealing surfaces on the pump housing 22. Also, the bracket 104 can be inexpensively and easily machined as a replacement part to the correct dimensions to match the dimensions of the recess 82 of the accumulator housing 34.

An annular disc groove 126 formed in the upper surface of the disc 112 adjacent the housing 34 communicates with the corresponding fuel passage 100. A pair of axial disc inlet passages 128 extend from the annular disc groove 126 to opposite opposite sides into communication with the pump chamber 124. A disk outlet passage 130 extending through the center of the disk 112 is aligned with a check valve recess 132 formed in the accumulator housing 34 adjacent the disk 112. A pump unit outlet passage 83 extends from the check valve recess 132 through the accumulator housing 34 into communication with the accumulator chamber 36c. A pump unit check valve 136 is disposed in the check valve recess 132 and is configured to sealingly engage the upper annular surface of the disk 112 surrounding the outlet passage 130 to prevent the flow of high pressure fuel from the chamber 36c when the pressure of the fuel in the chamber 36c is higher than the pressure of the fuel in the pump chamber 124, while allowing fuel flow from the chamber 124 into the accumulator 36c when the pressure in the pump chamber 124 exceeds the fuel pressure in the accumulator chamber 36c.

A corresponding recessed discharge passage 96 extending from the common fuel passage 90 communicates with an annular recessed gap 138 formed between the annular upper surface of the pump bracket 104 and the accumulator housing 34. A pump unit gap 140 formed between both the pump disk 112 and the bracket 104 and between the cylinder 116 and the bracket 104 is in continuous communication with the recessed gap 138. A bracket discharge passage 142 formed in the cylinder 116 extends radially outward from the central bore 118 into communication with the pump unit space 140 adjacent to the lower portion 108 of the bracket 104. An annular discharge groove 144 formed in the pump plunger 122 intermittently communicates with the discharge passage 142 during reciprocation of the pump plunger 122. Fuel leaking from the pump chamber 124 between the cylinder 116 and the plunger 122 collects in the discharge groove 144 and is intermittently discharged into the discharge passage 142. Fuel from the discharge passage 142 is continuously discharged through the pump unit space 140, the recess space 138 and the discharge intake outlet passage 96 into the common fuel supply passage 90.

As shown in Figures 5 and 6, the lower end of the pump plunger 122 extends through the lower wall 110 of the bracket 104 for engagement with a button 146 of a follower assembly 148. The button 146 includes an upper hemispherical seating surface for engagement with a complementary hemispherical surface formed on the lower end of the pump plunger 122. The follower assembly 148 also includes a follower housing 150 having a cylindrical outer surface mounted for reciprocating movement relative to corresponding cylindrical follower guide surfaces 152 formed on a portion of the vertical inner walls of the pump housing 22. The driver guide surfaces 152 are machined to ensure smooth sliding contact between the driver housing 150 and the pump housing 22 during reciprocation of the housing 150. A lower spring seat 154 disposed about the knob 146 and the lower end of the plunger 122 engages both the knob 146 and a retaining ring 156 disposed in an annular groove 157 formed on the plunger 122. A tension spring 158 disposed about the lower portion 108 of the bracket 104 engages at one end with a step 160 formed between the upper portion 106 and the lower portion 108 of the bracket 104. The opposite end of the tension spring 158 extends through the pump cavity 28 into engagement with the lower spring seat 154, thereby biasing the follower assembly 148 and the plunger 122 toward the camshaft 26. A roller 162 having a central bore 164 is disposed in an interior cavity 166 formed in the follower housing 150. The roller 162 is rotatably supported on the housing 150 by a pin 168 which extends through the bore 164 into openings 170 formed in the follower housing 150 on opposite sides of the cavity 166. Therefore, each roller 162 associated with each drive housing 150 is biased by the spring 158 against a corresponding cam 172 formed on the camshaft 26.

The cams 172 are disposed in the camshaft cavity 24 between a first opening 200 and a second opening 202 formed in the lower portion 223 of the pump housing 22. The camshaft 26 is secured to an engine shaft (not shown) by a Woodruff key 173 or any other conventional means for connecting two rotating shafts. The camshaft 26 rotates at half the engine speed so that each cam 172 rotates 360 degrees for every 720 degree rotation of the engine crankshaft. Each cam 172 includes at least one lobe 204 which places the associated pump plunger 122 in a forward or pumping stroke and a reverse stroke for each revolution of the camshaft. However, to provide, maintain and control the high fuel pressure in the storage chambers 36, it is advantageous to replenish fuel in the storage chambers 36 in synchronism with the removal of fuel from the storage chambers 36. To accomplish this series operation, the forward stroke number must be equal to the number of engine cylinders. In the six-cylinder engine of the preferred embodiment, two pump units 86 and 88 are each driven by a respective cam 172 provided with three lobes 204, so that the total number of lobes and hence the total forward stroke number is equal to the number of engine cylinders, i.e., six. In this way, each forward stroke of the pump plungers 122 directly corresponds in time to a delivery period associated with the fuel rail 16 and therefore to an injection period of an injector (not shown). Therefore, the lobes 204 are arranged around each cam 172 so that a pulse of fuel can be directed from the pump units 86 and 88 to the storage chambers 36 during the same period in which a pulse of fuel is withdrawn from the storage chambers 36 for delivery to the injectors by the distributor 16.

During operation of the pump 14, the pump control valves 18 and 19 are normally de-energized in an open position. Thus, during the return stroke of each pump plunger 122, fuel flows from the common fuel supply passage 90 through respective fuel supply side channels 78 and 80 into respective pump chambers 124. Likewise, during the pumping or forward stroke, each pump plunger 122 forces fuel from its respective pump chamber 124 through the fuel supply side passages 78 and 80 and corresponding pump control valves 18 and 19. However, when the fuel pressure in the accumulator chambers 36 falls below a predetermined minimum, the ECU 13 energizes the pump control valves 18 and 19 as needed at a predetermined point during a corresponding pumping stroke of the pump plungers 122, thereby closing the corresponding pump control valve 18, 19 and blocking the flow of fuel from the corresponding pump chamber 124. Further advancement of the pump plunger 122 pressurizes the fuel in the pump chamber 124 until the fuel pressure in the chamber 124 exceeds the fuel pressure in the accumulator chambers 36, causing the pump unit check valve 136 to unseat and allow fuel to flow from the pump chamber 124 into the accumulator chambers 36, thereby maintaining the fuel pressure in the accumulator 12 within the desired pressure range. The supply of fuel from chamber 124 to accumulator 12 ceases when pump plunger 122 completes its forward or pumping stroke. In this manner, pump 14 and associated pump control valves 18 and 19 are operated to control the effective displacement of each pump chamber 124 by providing a variable start of injection upon closure of a corresponding pump control valve 18, 19, while a constant end of injection occurs when pump plunger 122 reaches its top dead center or most advanced position. However, other forms of high pressure variable displacement pumps may be used to control accumulator pressure. Examples of such other variable displacement pumps are set forth in U.S. Patent No. 4,502,445 to Roca-Nierga et al. and in the co-pending patent application filed on the same date as the present application in the name of Yen et al. and entitled Variable Displacement High Pressure Pccmp for Common Rail Fuel Injection Systems, and assigned to the assignee of this invention. The entire disclosure of this application is incorporated herein by reference.

With reference to Figures 5 and 17a-27, the fuel distributor housing 44 of the distributor 16 is secured to the lower portion 23 of the pump housing 22 adjacent to the second opening 202. The fuel distributor housing 44 includes a rotor bore 214 extending axially through the housing 44 extends in axial alignment with the second opening 202 of the pump housing 22. An annular sealing recess 206 is formed in the distributor housing 44 at one end of the rotor bore 214 for receiving shaft seals 208 which prevent fuel leaking around the rotor 216 from entering the camshaft cavity 24. A rotor 216 is rotatably mounted in the rotor bore 214 and is connected at a first end to the camshaft 26 by a coupling 218. A second end of the rotor 216 terminates adjacent the inner surface of a recess 220 formed in the end of the distributor housing 44 adjacent to the rotor bore 214 (Figs. 5, 22 and 25). The recess 220 includes internal threads for engagement with the external threads of a discharge fitting 222 having a discharge opening 224 extending therethrough. Although the distributor housing 44 preferably extends from the pump housing 22, the housing 44 may be secured to the pump housing 22 so that the rotor 216 is perpendicular to the axis of the camshaft 26. Shown in schematic form in Fig. 17b. In this arrangement, the rotor 216 may be operatively connected to the camshaft 26 by gears 217.

The rotor 216 includes an axial supply passage 226 which extends axially inwardly along, but radially spaced from, the central axis of rotation of the rotor 216 from the second end of the rotor 216 and terminates at a point forward of the first end (Figs. 5 and 27). A plug 228 is threaded into the open end of the axial supply passage 226 adjacent the recess 220 to fluid-tightly close the passage 226 from the discharge passage 224. A radial feed passage 230 extends radially from the axial feed passage 226 into communication with the rotor bore 214. Six fuel receiving openings 231 and six corresponding fuel receiving passages 232 are formed in the distributor housing 44 and are evenly spaced around the circumference of the rotor bore 214 for sequential communication with the radial feed passage 230 during rotation of the rotor 216. A semi-annular balancing groove 234 formed in the rotor 216 extends approximately 75% or 272º of the circumference of the rotor 216. The balancing groove 234 terminates on both sides of the radial feed passage 230 so that when the feed passage 230 is aligned with one of the receiving passages 232, the remaining Pickup passages 232 communicate with the balance groove 234. Therefore, the fuel pressure in the pickup passages 232 communicating with the balance groove 234 is equalized prior to the start of each injection period. This equalization or relief of the initial fuel pressure in the pickup passages 232 and the corresponding downstream passages provides for controllable and predictable fuel metering from one injection period or engine cycle to the next. Further, an axial run-off passage 233 formed in the rotor 216 extends inwardly from the end of the rotor 216 adjacent the run-off fitting 222 into communication with a radial passage 235 extending radially inwardly from the balance groove 234. In this manner, the fuel in the equalization groove 234, and hence in the receiving passages 232 not communicating with the radial supply passage 230, is continuously connected to the fuel outlet which is maintained at a relatively constant low pressure. This maintains each receiving passage 232 at a relatively predictable, constant pressure so that pressurization of each receiving passage 232 begins at approximately the same pressure, thereby improving controllability and predictability of fuel metering. The opposite end of each receiving passage 232 communicates with a recess 236 formed in the end of the distributor housing 210. Each recess 236 has internal threads for engagement with complementary external threads on an outlet fitting 238. An axial injection bore 240 extends axially through each outlet fitting 238 into communication with a corresponding receiving passage 232. The receiving passages 232 are formed by inwardly drilling each recess 236 through the distributor housing 44 at an acute angle to the rotor axis. In this manner, each outlet fitting 238 fluid-tightly closes off the portion of the bore that extends radially outward from the fitting 238, thereby forming a fluid-tight connection between each receiving passage 232 and each injection bore 240. A radial receiving passage 242 formed in the rotor 216 and axially spaced from the radial supply channel 230 extends outwardly from the axial supply passage 226 into communication with an annular supply groove 244.

The portion of the present fuel supply system for supplying fuel from the storage chambers 36 to the supply groove 244 will now be described in detail. As shown in Fig. 5, fuel is delivered from the storage chamber 36a to the distributor housing 44 via the fuel supply passage 67 and a fuel supply tube 246. A fuel supply recess 248 formed in the open end of the fuel passage 67 includes a supply tube seat 250 for engagement with a supply tube head 252 formed on the end of the supply tube 246. The supply recess 248 has an internal thread for engagement with a complementary external thread formed on a substantially cylindrical supply tube fitting 254. The supply tube 246 extends through the tube fitting 254 so that one end of the tube fitting 254 rests against the tube head 252. Rotation of the tube fitting 254 with respect to the supply recess 248 and the fuel supply tube 246 forces the supply tube head 252 inwardly into sealing engagement with the tube seat 250, thereby creating a fluid-tight connection between the supply passage 67 and the supply tube 246. The supply tube 246 extends downwardly into the space between the distributor housing 44 and the cantilevered axial overhang 38 of the distributor housing 34 into a supply tube receiving recess 256 formed in the upper surface of the distributor housing 44. A cylindrical seal 258 formed at the end of the supply tube 246 is pressed radially outwardly against the surface of the receiving recess 256 to prevent leakage of fuel between the supply tube 246 and the receiving recess 256. An annular seal groove 260 formed in the recess 256 serves to receive a seal to prevent leakage of fuel from the recess 256 between the supply tube 246 and the housing 44. An annular supply tube drain groove 262 formed in the recess 256 between the seal groove 260 and the cylindrical seal 258 collects the fuel that leaks upward into the recess 256 between the supply tube 246 and the housing 44. A discharge passage 263 extends from the discharge groove 262 into connection with the discharge system of the first injection control valve 20.

An axial supply bore 264 extends axially outward from the transverse surface of the distributor housing 44 adjacent the second opening 202 of the pump housing 22 into communication with a first injection control valve cavity 270 formed in the distributor housing 44 for receiving a first injection control valve 20 (Fig. 24). The axial supply bore 264 continues axially outward from the first injection control valve cavity 270 into communication with a passage 266 extending from the recess 256. The open end of the transverse bore 264 includes a recess 268 which is fluid-tightly closed with a plug (not shown). A second injection control valve cavity 272 is formed in the distributor housing 44 adjacent to the first injection control valve cavity 270 so that the first and second injection control valve cavities 270 and 272 are located on opposite transverse sides of the rotor 216, respectively. A transverse feed bore 274 extending from one side of the distributor housing 44 above the rotor 216 fluidly communicates the first injection control valve cavity 270 with the second injection control valve cavity 272 (Figs. 21 and 23). The transverse feed bore 274 and the axial feed bore 264 are formed in the same horizontal plane so that they intersect the first injection control valve cavity 270 at adjacent points around the circumference of the cavity 270. The open end of the transverse supply bore 274 is fluid-tightly closed with a plug 275 (Fig. 23). A rotor supply bore 276 formed in the distributor housing 44 extends from one side of the housing 44 beneath the rotor 216 into communication with a first outlet passage 278 and a second outlet passage 280 extending from the first and second injection control valve cavities 270 and 272, respectively (Figs. 19, 23-26). The open end of the rotor supply bore 276 is fluid-tightly closed with an appropriately sized plug, similar to plug 277. A rotor channel 282 extends vertically from the rotor supply bore 276 upward into communication with the rotor bore 214. The supply channel 282 is formed by drilling upward through the bottom of the distributor housing 44. Therefore, the open end of the bore associated with the supply channel 282 is fluid-tightly closed with a plug (not shown).

The supply passage 282 and the rotor supply bore 276 are formed in a common vertical plane with the radial receiving passage 242 and the supply groove 244 so that the supply passage 282 is in constant communication with the supply groove 244 and the radial receiving passage 242 as the rotor 216 rotates. As a result, the fuel delivery to the axial supply passage 226 via the radial receiving passage 242, the supply groove 244, the supply passage 282, the rotor supply bore 276 and the first and second outlet passages 278 and 280 from the transverse bore 274 depends only on the position of the corresponding injection control valves 20 and 21. However, a two-way check valve is disposed in the rotor supply bore 276 to prevent fuel supplied from one of the injection control valve cavities 270 and 272 from flowing into the other injection control valve cavity. The first and second injection control valves 20 and 21, each operable to connect the axial supply passage 226 to the storage chamber 36a, may be of the three-way type shown in Fig. 23 and described in detail in a co-pending patent application filed March 19, 1993 in the name of Pataki et al., entitled Force Balanced Three-Way Solenoid Valve, and assigned to the assignee of this invention. The entire disclosure of that application is incorporated herein by reference.

The first and second injection control valves 20 and 21 are also operable to fluidly connect the axial supply passage 226 to a low pressure fuel outlet circuit generally designated 284 (Fig. 22). The outlet circuit 284 includes first and second axial outlet passages 286 and 288, respectively, which extend axially from the transverse surface of the distributor housing 44 adjacent to the pump housing 22 into communication with the first and second injection control valve cavities 270 and 272, respectively. The axial outlet passages 286 and 288 also extend axially from the respective cavities 270 and 272 into communication with the outlet passages 290 and 292, respectively (Fig. 22). The outlet passages 290 and 292 each extend inwardly at an angle to the axis of the rotor 216 into communication with an annular outlet collection groove 294 formed in the recess 220. A pair of outlet openings 296 and 298 formed in the innermost end of each outlet fitting 222 extend from the outlet collection groove 294 to the outlet opening 224 to direct fuel from the outlet collection groove 294 to a low pressure fuel outlet connected to the opposite end of the outlet fitting 222 (Fig. 5).

The drain circuit 284 further includes an axially extending drain passage 300 formed in the distributor housing 44 and communicating at one end with the seal recess 206 and at an opposite end with the drain passage 292 (Figs. 17a, 22 and 23). Therefore, fuel leaking through the gap between the rotor 216 and the distributor housing 44 into the seal recess 206 is directed to the drain. A vertical outlet passage 302 communicates at one end with a second valve recess 304 formed at the upper end of a valve cavity 272 and at a second end with the axial outlet passage 288. A first valve recess 306 communicates with a second valve recess 304 through two outlet passages 308 and 310 extending inwardly from respective recesses 306 and 304 (Figs. 20 and 23). Therefore, fuel leaking from the valve cavities 270 and 272 is collected in the recess 306 and 304, respectively, and directed to the outlet from the vertical outlet passage 302, the axial outlet passage 288, the outlet passage 292, the outlet opening 298, and the outlet opening 224.

Referring to Fig. 5, a safety valve 312, shown in schematic form, is disposed along the fuel delivery circuit in the delivery pipe 246 between the accumulator 12 and the injection control valve 20. During operation of the fuel pumping system, the injection control valve 20 may inadvertently become jammed or blocked in the open position, placing the accumulator 12 in constant fluid communication with the manifold 16. Therefore, high pressure fuel may flow from the accumulator 12 through the manifold 16 to the engine cylinders throughout the duration of each injection period. Thus, regardless of the engine throttle position, fuel is undesirably continuously supplied to the engine, possibly resulting in engine run-up. The safety valve 312 prevents such run-up by blocking the flow of fuel to the manifold 16 when the injection control valve 20 falsely remains in the open position. The safety valve 312 may be a pressure balanced, solenoid operated, two-way, two-position valve that completely blocks the flow of fuel through the supply tube 246. Alternatively, the safety valve 312 may be a pressure balanced, three-way valve, similar to the injection control valve 20, that is movable from an open position in which flow from the accumulator 12 to the manifold 16 is possible under normal operating conditions, to a drain position in which flow to the manifold 16 is blocked while the accumulator 12 is connected to a drain passage 314 via the supply tube 246. The safety valve 312 may be controlled by a signal from an ECU (not shown) indicating that the injection control valve 20 has not reached the closed position upon receipt of a closure signal. In addition, the safety control valve 312 can alternatively be arranged in the fuel delivery circuit between the injection control valve 20 and the distributor 16.

Reference is now made to an alternative embodiment of the present invention, shown in Fig. 28. In this embodiment, the same basic components are shown as were referred to in connection with the first embodiment of Figs. 2-6, namely a pump 401, an accumulator 402 and a distributor 404. However, unlike the previous embodiment, the pump assembly 400 of Fig. 28 includes a gear-type booster pump 406 disposed in a complementary cavity 408 located in the distributor housing 410. The function of the booster pump 406 is to ensure that the pump chambers 412 and 414 are filled with fuel during the downward stroke of the respective pump plungers 416 and 418. During certain operating conditions, such as high engine speeds, the downward stroke of the pump plunger 416 and 418 occurs at a rate exceeding the capacity of the normal engine "feed" pump to fill the respective pump chambers 412 and 414 with fuel.

To solve the problem associated with the pumps not always being completely filled or charged, the pressure booster pump 406 is provided to increase the pressure of the fuel supplied to the chambers 412 and 414. For example, the booster pump 406 can increase the supply pressure of fuel supplied to the pump chambers from a low value, e.g., 5 psi (about 34.47 KPa) to a significantly higher value, e.g., 200-300 psi (about 1.379 to 2.068 MPa). This significantly higher pressure generally causes the chambers 412 and 414 to be completely filled even during periods of maximum downward velocity of the respective pump plungers 416 and 418.

The pump 406 includes two meshing gears 420 and 422 housed in the cavity 408. The gear 422 is mounted on a shaft 424 that is coaxial with the drive shaft of the pump 401 and is connected thereto for drive rotation. The other end of the shaft 424 is connected to a distributor rotor 425 that functions similarly to the rotor 216 of the embodiment of Fig. 5. A spacer housing 426 is disposed between the pump housing 428 and the distributor housing 410 to facilitate mounting of the distributor and the pressure booster pump to the pump housing 428. A bearing journal 430 is provided in the spacer housing 426 for one end of the shaft 424. A fluid sealing ring 432 may be provided around one end of the drive shaft to maintain the separation of fuel in the pressure boost pump and the lubricating fluid in the drive shaft cavity 434 of the high pressure pump 401.

The high pressure fuel is stored in the accumulator 402 for delivery to the distributor 404 through a delivery tube 436. Although not shown in Fig. 28, passages are provided inside the distributor housing 410 which supply the axial supply passage 438 in the rotor 425 with high pressure fuel which is sequentially delivered to the individual engine cylinders in the manner previously described. Two solenoid-operated injection control valves 440 (only one of which is visible in Fig. 28) are provided for controlling the timing and amount of fuel injection into each engine cylinder by controlling the flow of fuel from the delivery tube 436 into the axial supply passage 438. The injection control valves 440 may also be of the three-way type shown in Fig. 23 and described in detail in a co-pending patent application. patent application filed March 19, 1993 in the name of Pataki et al., entitled Force Balanced Three-Way Solenoid Valve, and assigned to the assignee of this invention.

Another type of solenoid-operated injection control valve 440 is shown in Fig. 29. Two such valves 440 and 440' are shown in Fig. 29 as would be seen in a transverse cross-section of the manifold 404 taken along line 29-29 of Fig. 28. This type of valve is characterized by the provision of a "pin-in-sleeve" valve element which is force balanced but includes a high pressure valve seat 442 whose effective sealing area is significantly smaller than the outlet valve seat 444. When valve 440 is actuated, supply passage 446 is connected through valve seat 442 of the three-way valve to a supply bore 448, which in turn communicates with rotor receiving bore 450 through a connecting passage 452. The advantage of this type of valve is that the flow characteristics of the valve after opening can be made significantly different than the flow characteristics after closing. Also, a two-way check valve 453 is disposed in supply bore 448 to prevent fuel supplied from one of the injection control valve cavities from flowing into the other injection control valve cavity. This type of three-way control valve is also described in more detail in the co-pending patent application filed March 19, 1993 in the name of Pataki et al. filed, entitled Force Balanced Three-Way Solenoid Valve and assigned to the assignee of this invention.

Reference is now made to Fig. 30 which discloses a further embodiment of the present invention. In this embodiment, a single, solenoid-operated injection control valve 454 is provided instead of the two three-way valves of Fig. 23 or Fig. 29. In particular, the injection control valve 454 has its own valve housing 456 which contains a valve cavity 460 in which a three-way valve of the type shown in Fig. 29 is received. However, unlike the injection control valves of Figs. 23 and 29, the injection control valve 454 is arranged with the central axis of the valve cavity 460 parallel to the axis of rotation of the distributor rotors 462 of the distributor 464. High pressure fuel from the accumulator 466 is directed through a feed tube 468 to the valve cavity 460. When the solenoid 470 is actuated, the valve element 472 moves to the right in Fig. 30 to connect the feed tube 468 to the passage 474 which in turn directs the high pressure fuel through the passage 476 to the distributor bore 475.

Fig. 30 also shows a spacer housing 478 which differs from the spacer housing shown in Fig. 28 by the provision of a low pressure accumulator 480. The purpose of this additional accumulator is to enable the provision of an adequate volume of fuel to supply the pump chambers 482 and 484 of the high pressure pump 486 even during the period of highest retraction speed of the pump plungers 490 and 492. Without the low pressure accumulator 480, the gear pump would have to be larger to handle the high flow rate required during the period of highest retraction speed of the pump plungers 490 and 492. The fuel flow through the fuel pump assembly is as follows: Fuel is supplied from a fuel source, such as a fuel tank (not shown), to the assembly, to the gear pump 494, which is located in a separate gear pump housing 495. From the gear pump, the fuel is supplied through a first communication passage 496 (shown schematically in dashed lines) to the low pressure accumulator 480, and from the low pressure accumulator through a series of passages contained in the spacer housing 478, the pump housing 500, and the accumulator 466 to a supply passage 498 located in the high pressure accumulator 466. In particular, the fuel effluent from the low pressure accumulator 480 is supplied through a second delivery passage 502 to the pump housing 500.

Reference is now made to Fig. 31 which is a sectional view of the pump housing 500 taken along line 31-31 of Fig. 30. Fuel from the second delivery passage 502 is received in a horizontal passage 504 and directed upwardly through a vertical passage 506 which communicates through a storage delivery passage 508 with a supply passage of the accumulator 466 as shown in Fig. 32 which is a Section of the pump housing 500 and accumulator 466 taken along line 32-32 of Fig. 30. From the supply passage 498, fuel flows through passages 514 and 516 to the pump control valve recesses 510 and 512, respectively, as shown in Fig. 33, which is a partial sectional view of the accumulator 466 taken along line 33-33 of Fig. 30. Unlike the passages shown in Fig. 10e, the supply passage 498 is blocked at 518 (Figs. 32 and 33) so that fuel leakage returned to the supply passage 498 through passages 520 and 522 from the pump units shown in Fig. 30 does not mix with the fuel supplied to the pump control valves. Instead, as shown in Figs. 31, 32 and 33, the fuel is returned to the low pressure inlet of the gear pump 494 in the pump housing 495 through a series of passages designated 524, 526, 527 and passages not shown and formed in the spacer housing 487 and 495.

A series of spill passages are also provided in the injection control valve housing 456, the distributor housing 528, and the gear pump shafts 530 and 532. That is, these passages include a spill passage 534 extending radially through the valve housing 456 to direct fuel spills from the injection control valve 454 to an annular spill passage 536 formed in the upper surface of the distributor 464 which also collects leakage from the high pressure connection of the passages 474 and 476. A spill passage 538 extends inboard of the passage 536 and is connected to an annular cavity 539 formed around one end of the distributor rotor 462 which also collects fuel leakage between the rotor 462 and the distributor housing 528. The annular cavity 539 communicates with the inlet of the gear pump 494 through drain passages 541 and 543. The passage 541 also communicates with a drain cavity 544 which receives fuel leakage between the rotor 462 and the housing 528 through drain passages 546 and 548. Likewise, a drain passage 550 extends from an annular cavity 552 formed between lip seals 554 disposed about one end of the crankshaft 556 to drain fuel accumulated in the cavity 552 to a drain, not shown. In addition, two drains collect fuel from the annular cavity 552. passages 540 and 542 extending axially through gear pump shafts 530 and 532, respectively, direct fuel leakage between gear pump shafts 530 and 532 and spacer housing 478. Passage 542 directs fuel leakage to cavity 544 while passage 540 directs fuel leakage to cavity 539. A check valve 545 disposed in passage 540 is biased to prevent fuel leakage to the right in Fig. 30 until a low fluid pressure, e.g., 5 psi, is reached in passage 540. This arrangement prevents gear pump 494 from drawing air into its inlet from passage 550 and camshaft cavity 558.

Reference is now made to Figs. 34a and 34b which disclose two embodiments of the low pressure accumulator 480. Referring to Fig. 34a, the low pressure accumulator 480 includes a movable piston 560 slidably disposed within a cavity 562 extending through the spacer housing 478. Sealing plugs 564 are threaded into each end of the cavity 562 on opposite sides of the piston 560 for fluid-tight closure of the cavity 562. The piston 560 includes a first portion 566 slidably received in one of the sealing plugs 564 and a second portion 568 slidably and sealingly engaged with an inner wall of the housing 478 to divide the cavity 562 into a supply portion 570 and a discharge portion 572. A pressure regulating disk 574 disposed in the discharge portion 572 is biased to the left in Fig. 34a against an annular step 575 by a high pressure spring 576. A low pressure spring 578, seated at one end against the pressure regulating disk 574 and at a second end against the piston 560, biases the piston to the left in Fig. 34a. Fuel from the gear pump 494 (Fig. 30) enters a supply section 570 via a supply port (not shown) formed opposite an outlet port 580 connected to passages 502, 504, 506 and 508 which supply fuel to the high pressure fuel pump. The fuel flows through passages 582 and 583 which pass through the first portion 566 and acts on both sides of the first portion 566 and on an end surface of the second portion 568. As the pressure in the cavity 562 increases, the fuel pressure acts on the piston 560 and moves the piston 560 in Fig. 34a against the force of the low pressure spring 578 to create a fuel reservoir in the cavity 562. When the fuel demand of the high pressure pump exceeds the capacity of the gear pump, the spring 578 urges the piston 560 to the left, supplementing the fuel available from the gear pump. The arrangements of Figs. 34a and 34b also serve to regulate the pressure in the accumulator cavity 562. As the output of the gear pump increases, the higher fuel pressure urges the piston 560 against the pressure regulating disk 574, thereby urging the disk 574 to the right in Fig. 34a against the biasing pressure of the high pressure spring 576 until a left edge 584 of the second portion 568 is moved to the right side of a surface 586, allowing fuel to flow from the supply section 570 into the discharge section 572. Fuel in the discharge section 572 is returned to the inlet of the gear pump 494 via a discharge port 588 and return passages (not shown). Once the fuel pressure in the supply section 570 drops to a predetermined value, the high pressure spring 576 urges the piston 560 to the left and fluid-tightly closes the supply section 570 from the discharge section 572. In this way, the accumulator 484 maintains an adequate supply of fuel to the pump chambers 482 and 484 of the high pressure pump 486 even during the period of the highest retraction speed of the pump plungers 490 and 492 (Fig. 30).

Fig. 34b shows a second embodiment of a low pressure accumulator 480 with a movable piston 590 which is arranged in a cavity 592 which is formed in one side of the spacer housing 478 and is closed fluid-tight by a sealing plug 593. Supply fuel passes into and out of the supply section 594 via passages 596 and 598. When the pressure in the cavity 592 increases, the piston 590 is moved to the right in Fig. 34b against the biasing pressure of a low pressure spring 600. When the fuel pressure increases to a predetermined value, the piston 590 comes into contact with the pressure regulating disk 602 and moves the disk 602 to the right against the biasing pressure of a high pressure spring 604 so that supply fuel can flow out through the passage 606. When the supply fuel pressure drops, the spring 604 the disc 602 returns to its closed position against a step 608.

Referring to Figure 35, an alternative hydromechanical embodiment of the present invention is disclosed which is similar to the previously discussed embodiments in which a high pressure pump unit 700 supplies high pressure fuel to an accumulator 702 for sequential delivery to a plurality of injector nozzles, one of which is shown at 704, via a fuel distributor 706 including a rotor 708 which rotates to sequentially deliver fuel from the supply channels 710 formed in the rotor 708 to receiving passages 712 formed in the distributor housing 713. However, unlike the previous embodiments, the rotor 708 is mounted for axial displacement under the influence of an engine speed sensing centrifugal device 714 at one end and by a spring element 716 at another end, the tension of which is adjustable depending on the rotation of a cam 718 which may be controlled by the throttle position and/or a speed controller. The supply passages 710 include a pilot injection passage 720 leading to the supply passages 710 to perform pilot injection and a generally triangular shaped main injection passage 722. The shape of the passage 722, which is aligned with the receiving passages 712 upon further rotation of the rotor 708, changes in the axial direction of the rotor 708 so that the amount of fuel injected by the corresponding fuel injector is changed in accordance with the axial position of the rotor 708. To change the timing of each injection event performed by the system, a "phasing mechanism" 724 may be provided which advances or retards the rotor 708 with respect to the instantaneous position of the camshaft. Such a mechanism may be responsive to a mechanical, electrical or fluid signal to adjust the angular position of the rotor 708 relative to the engine camshaft.

Referring now to Fig. 36, another embodiment of the present invention is shown which is similar to the embodiment shown in Fig. 1, except that a rotary pump 750 is used instead of of the high pressure in-line pump 14 disclosed in Fig. 1. The rotary pump 750 includes pump plungers 752 mounted for reciprocating movement in pump chambers 754 formed in a portion of the drive shaft 756 which forms a rotatable pump housing. Alternatively, the pump chambers may be formed in a rotatable pump housing separate from the drive shaft 756 but adapted for rotation therewith. Preferably, the drive shaft 756 is also used to drive the distributor 758 which may be formed in the drive shaft 756 or may be formed as a separate rotatable assembly driven by the shaft 756. The distributor 758 functions in the same manner as the distributor 16 of Fig. 5.

A cam ring 760 through which the drive shaft 756 passes includes an inner annular cam surface 762 against which the pump plungers 752 are biased by, for example, tension springs (not shown). In this way, as the drive shaft 756 rotates, the pump plungers 752 are rotated with respect to the cam surface 762 which, depending on the contour of the cam surface 762, alternately presses the plungers 752 inward and causes the plungers 752 to move outward. The pump chambers 754 communicate with a common central cavity 764 which is continuously connected to a pump control valve 766, for example through an axial passage 768, a radial passage 770, an annular groove 772 and a connecting passage 774 formed in a pump housing (not shown).

Although not shown, the pump housing may be stationary and the cam ring 760 may be arranged for rotation with the drive shaft 756. The radially directed pump chamber may be arranged radially inward of the cam ring, as shown in Figure 36, or the pump chambers may be arranged radially outward of the cam surface. Regardless of the cam ring embodiment used, the rotary pump of Figure 36 may be integrated into the modular pump assemblies of the present invention, as disclosed in Figures 5, 28, and 30.

The operation of the embodiment disclosed in Fig. 36 is essentially the same as the embodiment of Fig. 1, except that the rotary pump 750 operates to simultaneously move the pump plungers 752 radially inward and outward during rotation of the drive shaft 756. When the pump valve 766 is open, fuel from a fuel supply (not shown) can flow through the pump control valve 766 into the pump chambers 754 upon the outward stroke of the pump plunger 752. The fuel is forced out through the pump control valve 766 back to the supply upon the inward movement of the pump plungers 752 as long as the pump control valve 766 is in the open position. When fuel delivery to the accumulator is desired, the pump control valve 766 is moved to the closed position during the inward stroke of the pump plunger 752 to block fuel flow to the supply so that the high pressure fuel can be delivered from the common center cavity 764 to the accumulator 776. This embodiment of the present invention is particularly advantageous in providing a highly compact, low cost pumping system that can be easily adapted for use with small engines subject to severe size, weight and price constraints. It should also be noted that only one pump control valve is required for a plurality of pump plungers, thereby simplifying the assembly and control system.

Referring now to Figures 37 and 38, an alternative embodiment of the fuel rail used in the fuel system of the present invention is disclosed. In particular, the rail manifold 780 includes a rail manifold housing 782 which includes rail manifold or injection valves 784 which are actuated by a rotatable camshaft 786 for delivering pressurized fuel through corresponding delivery valves 788 to corresponding engine cylinders (not shown). The rail manifold housing 782 includes a large cylindrical recess 790 in one end of the housing 782 for receiving the rotating camshaft 786. A seal 792 is provided between the outer annular surface of the camshaft 786 and the rail manifold housing 782 to prevent fuel from leaking between the camshaft 786 and the housing 782. while allowing rotation of the camshaft 786. The camshaft 786 includes an end surface 794 having a cam 796 formed thereon for actuating the injection rail valves 784 during rotation of the camshaft 786. The cam 796 is disposed on the outer radial portion of the end surface 794 for sequential contact with the injection rail valves 784.

The distributor housing 782 further includes a plurality of valve cavities 798 extending axially along the axis of rotation of the camshaft 786, perpendicular to the end surface 794. The valve cavities 798 are evenly spaced in a circular array as shown in Fig. 38 and extend from the inner end of the cylindrical recess 790. A supply inlet passage 800 is formed in the distributor housing 782 and is fluidly connected to one end of the injection control valve 20 of Fig. 1. The opposite end of the supply inlet passage 800 is connected to a common supply chamber 802 which is fluidly connected to each of the valve cavities 798. A corresponding fuel injection outlet passage 804 extends radially outward from each valve cavity 798 through the housing 782 for delivering high pressure fuel to corresponding fuel injection lines 806 leading to corresponding engine cylinders. The corresponding spring-loaded delivery valve 788 is disposed in each fuel injection line 806 to prevent the flow of fuel from each fuel injection line 806 back through the manifold 780.

The injection rail valves 784 are each of the plunger type with a slide valve element 808 disposed for reciprocating movement in a respective valve cavity 798. Each slide valve element 808 extends at one end into the inner end of the recess 790 adjacent the end face 794 of the camshaft 786 so that it is positioned for engagement with the cam 796 during rotation of the camshaft 786. The opposite end of each slide valve element 808 extends into its respective valve cavity 798 and beyond the connections of the fuel injection outlet passages 804 and supply chamber 802 to the valve cavity 798. A tension spring 810 is disposed in a cavity 811 defined by the opposite end of the sliding valve member 808 and a closed end of each valve cavity 798 to bias the sliding valve member 808 toward the camshaft 786 and into engagement with the end surface 794.

Each slide valve element 808 also includes a cylindrical surface 812 that is dimensioned to form a close sliding fit with the inner surface of the valve cavity 798, thereby creating a fluid seal between the adjacent surfaces that prevents fuel from leaking from the outlet passage 804 and the supply inlet passage 800 when the surface 812 covers or blocks these passages. The slide valve element 808 also includes an annular groove 814 formed in its outer surface to create a surface 812 at one end of the element 808. The annular groove 814 is formed along the valve element 808 so that it is brought into communication with the common supply chamber 802 and the fuel injection outlet passage 804 when the corresponding slide valve element is moved inwardly by the cam 796 against the biasing force of the spring 810.

The operation of the fuel rail of Fig. 37 will now be discussed in accordance with its use in the fuel pumping system of the present invention. As the camshaft 786 rotates, the cam 796 sequentially engages the slide valve members 808 of the injection line valves 784, moving a corresponding slide valve member 808 to the right as viewed in Fig. 37 against the bias of the spring 810. In this manner, the annular groove 814 moves into communication with the common supply chamber 802 and the fuel injection outlet passage 804, thereby placing the injection line valve 784 in an open position in which the supply inlet passage 800 is fluidly connected to a corresponding injection line 806. As the camshaft 786 continues to rotate, the cam 796 passes the end of the sliding valve element 808 so that the sliding valve element 808 can return to a closed position under the force of the tension spring 810, with the surface 812 blocking the flow between the common supply chamber 802 and the fuel injection outlet passage 804. The opening and closing of each injection line valve 784 defines a corresponding possible injection period or opportunity phase in which at which injection can occur as determined by the operation of the injection control valve 20 shown in Fig. 1. However, at any given time during rotation of the camshaft 786, only one injection line valve 784 is in an open position defining the injection period. The injection control valve 20 opens and closes subsequently during each injection period to define an injection event during which high pressure fuel is delivered from the high pressure accumulator 12 via the supply inlet passage 800, the common supply chamber 802, via a corresponding injection line valve 784 into the outlet passage 804 and a corresponding injection line 806 for delivery to a corresponding injector nozzle assembly 11 and the associated engine cylinder (not shown). The injector valve 784 also includes a balance passage 816 extending from one end of the slide valve member 808 to the opposite end so that the recess 790 is communicated with a spring cavity 811. In this manner, any pressure developed in the recess 790 and spring cavity 811 due to fuel leakage between the slide valve member 808 and the distributor housing 782 can be equalized so that movement of the slide valve member 808 is possible. Although not shown, a drain passage may also be used to connect the spring cavity 811 and/or the recess 790 to a low pressure fuel drain. Alternatively, the spring cavity 811 and recess 790 may be filled with lubricating oil via a passage (not shown) communicating with the recess 790. In addition, other forms of distributors may be used in the present fuel system, including the distributors disclosed in commonly assigned U.S. patent application Serial No. 117,697, entitled Distributor for High Pressure Fuel Injection System, which is incorporated herein by reference.

Fig. 39 and 40 show two further embodiments of the high pressure pump assembly of the present invention as shown in Fig. 6. Components of these embodiments which are the same as those disclosed in Fig. 6 are provided with the same reference numerals. Both embodiments of Fig. 39 and 40 advantageously reduce the number of components of the assembly and the complexity of manufacturing lation process, thereby advantageously reducing the cost of the entire system. Furthermore, these embodiments reduce the possibility of fuel leakage from the pump chamber by reducing the number of sealed joints exposed to high pressure fuel.

As shown in Figs. 39 and 40, these embodiments achieve the above advantages by avoiding the use of the sealing washer 112 of the embodiment shown in Fig. 6. The embodiment of Fig. 39 includes a one-piece pump cylinder 820 having an inner end 822 that is press-fitted to the accumulator housing or pump head 34 under the force of the bracket 104. The pump unit check valve 824 extends into a pump outlet passage 826 that extends through the inner end 822 along the central axis of the pump chamber 828. The pump unit check valve 824 is configured for sealing engagement with a check valve seat 829 formed on the upper annular surface of the pump cylinder 820 surrounding the pump discharge passage 826 to prevent the flow of high pressure fuel from the storage chamber 36c when the pressure of the fuel in the chamber 36c is higher than the pressure of the fuel in the pump chamber 828, while allowing fuel flow from the chamber 828 into the storage chamber 36c when the pressure in the pump chamber 828 exceeds the fuel pressure in the storage chamber 36c. The check valve 824 is biased against the check valve seat 829 to a closed position by a bias spring 830 disposed in a discharge passage 832. A spring guide pin 834 extends to guide the spring 830 from the storage chamber 36c into the discharge passage 832 while providing a seating surface for the spring 830. The pump cylinder 820 also includes two pump inlet passages 836 extending from the pump chamber 828 into communication with an annular groove 838 formed in the upper surface of the pump cylinder 820. As previously described in more detail with reference to Fig. 6, the annular groove 838 fluidly communicates with a pump control valve 18, 19 through a corresponding fuel passage 840 and a fuel supply side channel 842. The operation of this embodiment is in essentially the same as that previously described with reference to Fig. 6.

Referring now to Fig. 40, another embodiment of the pump assembly includes a pump cylinder 844 disposed in seating with the pump head 34 such that the pump chamber 846 is positioned immediately adjacent to the pump head 34. The pump head 34 extends over the pump chamber 846 to form at least a partial end wall 848 of the pump chamber 846. In this embodiment, no pump inlet and outlet passages are formed in the pump cylinder 844 as the pump inlet and outlet passages 850 and 852, respectively, are formed entirely in the pump head 34. A check valve 854 is disposed in the outlet passage 852 for seating against a check valve seat 856 formed annularly around the outlet passage 852. A check valve assembly cavity 858 extends downwardly from the upper surface of the pump head 34 into communication with the pump outlet passage 852 to allow for easy installation of the check valve 854 and its associated spring 860 and guide pin 862. A sealing plug 864 is threaded into the check valve assembly cavity 858 to seal the cavity 858 while providing support for the spring 860 and guide pin 862. Both embodiments shown in Figures 39 and 40 advantageously create only a single high pressure connection between the inner end of each pump cylinder and the adjacent pump head. This design minimizes the amount of fuel leakage and reduces the time and cost involved in forming "metal to metal" sealing surfaces, ensuring effective high pressure operation of the pump at reduced cost.

Reference is now made to Figs. 41 to 43 which disclose another embodiment of the present invention. This embodiment is substantially the same as the previously discussed embodiment shown in Fig. 30 with respect to the single solenoid-operated three-way injection control valve 454, the distributor 464, the gear pump 494 and the lower portion of the high pressure pump assembly 486. However, in this embodiment, an accumulator housing or pumps head 870 is integrated into the upper portion of the high pressure pump assembly 486 so that the overall height of the fuel pump assembly is minimized. In particular, pump chambers 872 and 874 are formed directly in the accumulator housing 870. The pump chambers 872 and 874 are formed along a respective radial pump axis which extends through outwardly opening pump cavities 876, 878 in which the pump units 880 and 882 are received. Pump plungers 884, 886 extend into the respective pump chambers 872 and 874 for reciprocating movement during rotation of the drive shaft 888. The pump chambers 872 and 874 are formed by the respective pump cylinders 890 and 892 which are integrally formed with the accumulator housing/pump head 870. Pump cylinders 890 and 892, which are integrally formed with accumulator housing 870, each extend into the interior of the respective pump cavities 876, 878 to support pump plungers 884, 886. Respective annular spring recesses 894 and 896 are formed around the respective pump cylinders 890, 892 to receive and support one end of the respective tension springs 898 and 900. The accumulator housing/pump head 870 also includes two pump valve recesses 902 and 904 formed in a side wall 906 and extending transversely into the housing to receive the pump control valves 18, 19. A corresponding cavity 908, 910 extends from each pump valve recess 902, 904, respectively, to an opposite side wall 912 laterally through the housing 870 for receiving a corresponding control valve element 914 (Fig. 43) of a corresponding pump control valve 18, 19. Each valve cavity 908, 910 is disposed axially along the housing 870 directly above the corresponding pump chamber 872, 874 so that the pump chambers 872, 874 open directly into the corresponding valve cavities 908, 910.

As shown in Figs. 41 and 42, annular grooves 916, 918 are formed in respective valve cavities 908, 910 transversely between the respective pump chamber 872, 874 and the side wall 912. A common axial discharge passage 920 extends axially through the housing 870 to connect the annular grooves 916 and 918. The common axial discharge passage 920 extends axially from the valve cavity 910 and intersects a transverse passage 922 formed by the Side wall 912 extends transversely through a portion of the accumulator housing 870. The open ends of the delivery passage 920 and transverse passage 922 are fluid-tightly closed by plugs 920a and 922a disposed in a recess formed in the open end. The accumulator housing 870 also includes two accumulator chambers 924 and 926 extending axially into the housing from an end wall 928. A corresponding axial passage 930, 932 connects each accumulator chamber 924, 926 to the transverse passage 922. As shown in Fig. 43, the accumulator housing 870 also includes a corresponding supply passage 934 associated with each pump control valve 18, 19. In general, pump control valves 18 and 19 are each preferably a solenoid-operated valve assembly similar to that disclosed in commonly assigned U.S. Patent No. 4,905,960 to Barnhart. The mounting arrangement of pump control valves 18 and 19 in pump head 870 is similar in construction. Only the differences in pump control valve 18 will be described below. In this particular application, pump control valve 18 includes a spring housing 936 disposed between a solenoid housing 938 and a valve seat member 940. Valve seat member 940 is disposed in a compressive, fluid-sealing, abutting relationship between spring housing 936 and an annular seating surface 942 formed on accumulator housing 870 about valve cavity 908. Valve seat member 940 extends radially inward to form an annular valve seat 944 surrounding valve cavity 908. The pump control valve 18 also includes a valve element 946 mounted for reciprocating movement within the valve cavity 908 to control the flow of fuel to and from the pump chamber 872. The valve element 946 has an annular conical surface 948 for engaging the valve seat 944 when the valve element 946 is moved to a closed position. An armature 950 is connected to one end of the valve element 946 adjacent the solenoid coil assembly 952 and is drawn toward the solenoid coil assembly 952 when the coil assembly is energized. A valve biasing spring 954 is disposed within an annular cavity 956 formed in the spring housing 936 to bias the conical surface 948 of the valve element 946 away from the valve seat 944 to an open position. The spring housing 936 is pivotable relative to the inner surface of the pump valve recess 902. to form an annular gap 958 in communication with the supply passage 934. The valve seat member 940 includes radial passages 960 in communication with the annular gap 958. The valve element 946 is arranged relative to the valve seat member 940 to form a first annular passage 962 in communication with the radial passages 960 on one side of the valve seat 944. On the opposite side of the valve seat 944, the valve element 946 is arranged relative to the annular inner surface of the valve cavity 908 to form a second annular passage 964 which communicates at one end with the first annular passage 962 when the valve element 946 is in the open position and at an opposite end with the pumping chamber 872.

As shown in Figure 43, the valve element 946 of the pump control valve 18 also includes a pump outlet passage 966 that connects the pumping chamber 872 to a check valve cavity 968 formed centrally in the valve element 946. A spring loaded check valve 970 is disposed in the check valve cavity 968 and is biased by a check valve spring 972 against a check valve seat 974 formed on the annular inner surface of the valve element 946 in the cavity 968. A spring guide pin 976 is also disposed in the check valve cavity 968 and is secured to the valve element 946 by an internal snap ring 978. Therefore, the check valve assembly, which includes the check valve 907, the check valve spring 972, and the spring guide pin 976, reciprocates with the valve element 946 during operation of the pump control valve 18. The open end of each valve cavity 908, 910 is fluid-tightly closed by a plug 980 which is threaded into a recess formed in the open end. A valve stop 982 is threaded to the plug 980 to provide a support for the outer annular end of the valve element 946 when the valve element 946 is moved to the open position by the tension spring 954. The valve stop 982 includes an inner extension 983 as a support for the guide pin 976. By rotating the valve stop 982 relative to the plug 980, the transverse position of the valve stop 982 relative to the valve element 946 and thus the valve stroke of the valve element 946 can be adjusted.

The valve element 946 further includes radial passages 984 arranged to allow fluid communication between the check valve cavity 968 and the annular groove 916. The check valve seat 974 is arranged along the check valve cavity 968 between the pump outlet passage 966 and the radial passage 984 so that the check valve 970 can prevent the backflow of high pressure fuel from the storage chambers 924, 926 when in the closed position while allowing high pressure fuel to flow from the pump chambers 872, 874 to the storage chambers 924, 926 when the valve element 946 moves to the closed position. The accumulator housing 870 also includes a discharge passage 986 that extends from the valve cavity 908 adjacent the valve stop 982 to a low pressure discharge (not shown).

The pump arrangement of Figs. 41-43 is particularly advantageous in several respects. Firstly, by forming the pump cylinders 890, 892 integrally with the pump head/accumulator housing 870 and mounting the pump control valves 18, 19 in the side of the accumulator so that they pass across the accumulator housing 870. The accumulator housing 870 can be moved closer to the drive shaft 888, thereby providing a more integrated, compact and lightweight pump arrangement. As shown in Fig. 41, this compact arrangement allows for adjacent placement of the injection control valve 454 between an axial overhang 987 of the accumulator housing 870 and the manifold 464. Instead of a vertical feed tube connecting the accumulator to the injection control valve as shown in the previous embodiments, a feed tube 989 is connected at one end to a plug 991 disposed in the open end of the accumulator chamber 926 and forms a loop in the connection with the side wall of the housing containing the injection control valve 454. Second, this integrated arrangement reduces the volume of high pressure fuel trapped in the high pressure passages during a pump delivery stroke because the pumping chambers are relocated immediately adjacent to the valve chambers and valve seats. This reduction in trapped volume results in increased pumping performance during each stroke of the high pressure pump because a larger part of the total volume of fuel subjected to very high pressure is actually transferred to the accumulator. Therefore, for a given size of the fuel pump assembly, the power of the engine can be increased, since less energy is consumed by the high-pressure pump in pumping the same amount of fuel into the accumulator compared to a similar installation without this feature. Thirdly, this design reduces the number of high-pressure connections between the pump chamber and the accumulator chambers, since the pump chamber is relocated inside the accumulator housing.

Referring now to Figures 44 and 45, another embodiment of the present invention is illustrated. Essentially, this embodiment discloses a novel pump assembly having a pump head 990, two pump units 992 and 993, and corresponding pressure balanced pump control valves 994 and 997. The pump units 992 and 993 and the associated pump control valves 994 and 997 are similar in construction and therefore only the pump unit 992 and pump control valve 994 will be discussed hereinafter. Although not shown, the fuel pump assembly 988 may be used with or mounted to the same components of the fuel pumping systems disclosed in Figures 5, 28 and 30, including the solenoid-operated three-way injection control valve or valves, the manifold and the lower portion of the high pressure pump assembly. As shown in Fig. 44, the pump unit 992 includes a pump cylinder 995 held in a pump recess 996 by a pump bracket 998 having an external thread for engagement with complementary threads formed on the annular inner surface of a bore 1000 formed in the outer end of the recess 996. The pump unit 992 also includes a pump chamber 1002 formed in the cylinder 995 and a pump plunger 1004 arranged for reciprocating movement in the pump chamber 1002 in response to rotation of the drive shaft (not shown). The pump cylinder 995 has an inner end 1006 arranged in contact with the pump head 990. A pump unit outlet passage 1008 extends from the pump chamber 1002 through the inner end 1006. A drain passage 1010 is provided in the pump head 990 for connecting the outlet drain passage 1008 with a storage chamber 1012. A pump unit check valve assembly 1014 is formed in the storage chamber 1012, the drain passage 1010 and the pump unit outlet passage 1008. The check valve assembly 1014 includes a check valve element 1016, a tension spring 1018 and a guide pin 1020. The check valve element 1016 is biased by the spring 1018 into engagement with an annular valve seat 1022 formed on the pump cylinder 995 about the outlet passage 1008 so that fuel flow from the storage chamber 1012 into the pump chamber 1002 is prevented while fuel flow from the pump chamber 1002 into the storage chamber 1012 is permitted when the fuel pressure in the chamber 1002 is greater than the fuel pressure in the chamber 1012. A cover plate 1024 and a sealing ring 1026 are disposed about the annular seat 1022 between the pump cylinder 995 and the pump head 990 to prevent leakage of high pressure fuel between these components. Alternatively, the cover plate 1024 and sealing ring 1026 may be omitted to form a metal-to-metal connection between the pump cylinder 995 and the pump head 990. An outer annular groove 1028 is formed between the pump cylinder 995 and the pump head 990 to receive high pressure fuel leaking through the sealed connection formed by either the cover plate 1024 and sealing ring 1026 or the metal-to-metal interface. A drain connection passage 1030 extends from the annular groove 1028 into communication with a combined outlet passage 1032 to direct the leaked fuel from the annular groove 1028 to the drain via a main drain passage 1034 formed in the pump housing. A similar outlet connection passage (not shown) associated with the pump unit 993 is connected to the main passage 1034.

A lubrication flow passage 1036 extends from the annular groove 1028 through the pump cylinder 995 into communication with an annular lubrication channel 1038 formed in the cylinder 995 around the chamber 1002. First and second annular lubrication grooves 1040 and 1042, respectively, are formed in the plunger 1004 and connected by a transverse passage 1044. During reciprocation of the plunger ger piston 1004 in the chamber 1002, the first and second annular lubrication grooves 1040, 1042 are intermittently connected to the annular lubrication channel 1038. In this way, low pressure fuel from the annular groove 1028 is used to lubricate the plunger piston 1004, thereby reducing friction between the plunger piston 1004 and the inner surface of the pump cylinder 995 and thus minimizing abrasion, scuffing and scoring due to wear of the contact surfaces.

A valve cavity 1046 extends diametrically through the pump cylinder 995 so as to intersect the inner end of the pump chamber 1002 and the outer end of the outlet passage 1008. The valve cavity 1046 also extends through the pump head 990 into communication with a plug recess 1048 at one end and a spring chamber 1050 at the opposite end. The open end of the valve cavity 1046 adjacent the recess 1048 is fluid-tightly closed by a plug 1052 which is threadedly secured to the pump head 990 in the recess 1048. The pressure balanced pump control valve 994 includes a valve actuator 1054 attached to one side of the pump head 990 and a control valve element 1056 mounted for reciprocating movement within the valve cavity 1046. The control valve element 1056 has an annular valve surface 1058 for abutting against an annular valve seat 1060 formed on the pump cylinder 995 around the valve cavity 1046 when the pressure balanced pump control valve 994 is in a closed position. A tension spring 1059 is disposed within the spring chamber 1050 to bias the control valve element 1056 to an open position. Delivery of fuel to the pump chamber 1002 is via a main supply passage 1062 formed in the pump housing, a communication passage 1064 formed in a lower portion of the pump head 990, and a transverse supply passage 1066 extending longitudinally through the pump head 990 to communicate the spring chamber 1050 of a pump control valve 994 with an adjacent pump control valve, as shown in Fig. 45. An annular channel 1067 is formed in the pump head 990 around the pump recess 996 adjacent the valve cavity 1046. An annular gap 1068 formed between the control valve element 1056 and the inner surface of the valve cavity 1046 connects the spring chamber 1050 to the annular channel 1067. At the opposite end of the valve cavity 1046, the annular channel 1067 is connected to the chamber 1002 through an annular gap 1070 formed between the control valve member 1056 and the inner surface of the valve cavity 1046. The annular valve seat 1060 is formed along the annular gap 1070 between the annular channel 1067 and the chamber 1002. In this way, the annular valve surface 1058 can be moved into and out of engagement with the annular valve seat 1060 to control the flow of fuel into and out of the pump chamber 1002.

The pressure balanced pump control valve 994 may be any conventional solenoid-operated, pressure balanced two-way valve that can be adapted for use in this design. The control valve element 1056 of the pressure balanced pump control valve 994 is pressure balanced in the closed position because the fluid pressure forces resulting from the high pressure fluid acting on the control valve element 1056 in one direction, i.e., to the right in Fig. 44, are equal to the fluid pressure forces resulting from the high pressure fluid acting on the control valve element 1056 in the opposite direction, i.e., to the right. to the left in Fig. 44, because the effective cross-sectional area of the valve seat 1060 which remains subject to the fluid pressure prevailing in the pump chamber is equal to the effective cross-sectional area defined in the portion of the valve element 1056 received in the pump cylinder on the right side of the pump chamber 1002, so that the control valve element 1056 which causes the rightward forces is equal to the surface area of the control valve element 1056 which causes the leftward forces.

During operation, fuel is delivered from a supply pump (not shown) into the spring chamber 1050 through the main supply passage 1062, the connecting passage 1064 and the transverse feed passage 1066. Fuel flows from the spring chamber 1050 through the annular gap 1068 surrounding the control valve element 1056, the annular channel 1067 surrounding the cylinder 995, and into an annular gap 1070 adjacent the annular valve seat 1060. When the pressure balanced pump control valve 994 is in the off position, open position, fuel flows between the annular valve seat 1060 and the annular valve surface 1058 into the pump chamber 1002. As the pump plunger 1004 reciprocates, fuel is supplied into and pumped out of the pump chamber 1002 via these supply passages. When fuel delivery to the storage chamber 1012 is required, the valve actuator 1054 of the pump control valve 994 is energized during the forward movement of the pump plunger 1004 so that the valve element 1056 is moved to the right in Fig. 44, thereby engaging the annular valve surface 1058 with the annular valve seat 1060. This blocks the flow of fuel through the annular gap 1070 so that the pump plunger 1004 can compress and pressurize the fuel remaining in the pump chamber 1002. Upon reaching a pressure value higher than the fuel pressure value in the storage chamber 1012, the fuel in the pump chamber 1002 opens the check valve element 1016 and flows through the outlet port 1008 and the drain port 1010 into the storage chamber 1012. Depending on the control scheme used, at some point during the forward or rearward movement of the pump plunger 1004, the pressure balanced pump control valve 994 is de-energized so that the check valve element 1016 can move to an open position under the force of the tension spring 1059. The advantage of using a pressure balanced valve is that there is more room to open and close the pump control valve. In particular, it becomes readily possible to terminate the effective pumping stroke of the pump plunger 1004 at any point in the forward stroke without generating very high spring or electromagnetic forces that would be required when using an unbalanced valve design.

Reference is now made to Fig. 46 which shows another embodiment of the present invention which is the same as the embodiment of Figs. 44 and 45 except that a pump head 1072 does not include storage chambers for collecting a quantity of fuel. As will be explained in more detail below in connection with the embodiment of Figs. 52 and 53, the pump head 1072 includes only a single common discharge passage 1074 for receiving fuel. of one or more pump chambers 1002. One end of the common delivery passage 1074 is connected to a remotely mounted accumulator positioned at a distance from the fuel pump assembly as shown in Fig. 52. This arrangement provides a more compact fuel pump assembly while at the same time allowing the high pressure accumulator to be mounted at a more convenient and advantageous location on the engine.

Fig. 47 shows another embodiment of the fuel pump assembly of the present invention, which is the same as the embodiments disclosed in Figs. 5, 28 and 30, except that a pressure balanced pump control valve 1076 is used. The pressure balanced pump control valve 1076 may be any conventional pressure balanced solenoid operated two-way valve. A pump control valve cavity 1080 extends upwardly from a valve recess 1082 formed in a lower surface of the accumulator housing 1078. The valve cavity 1080 opens into a plug recess 1084 which is fluid-tightly closed by a plug 1086. The plug 1086 terminates in front of the end wall of the recess 1084 to form a chamber 1088. The pump control valve 1076 includes a control valve element 1090 that extends through the valve cavity 1080 and terminates at one end in the chamber 1088. An annular valve seat 1092 formed around the valve cavity 1080 adjacent the chamber 1088 is configured to bear against an annular valve surface 1094 formed on the control valve element 1090. An annular recess 1096 may be formed in the valve cavity 1080 adjacent the control valve element 1090 between the valve seat 1092 and the valve recess 1082. An annular channel 1098 formed between the control valve element 1090 and the inner wall of the valve cavity 1080 fluidly connects the chamber 1088 to the annular recess 1096 when the control valve 1076 is in the open position.

The fuel supply passages formed in the accumulator housing 1078 are substantially the same as those disclosed in Fig. 5-101, except for the following changes. First, the connecting passages 92 and 94 shown in Fig. 10e and force ure from the common fuel supply passage 90 to both pump control valves, from each chamber 1088 downward into communication with the passage 90, rather than upward from the pump control valve recess 1082 as suggested in the embodiment of Figs. 5 and 10e. Also, the storage chamber 36a is necessarily shorter in length so that it terminates before the plug recess 1084. The operation of the embodiment of Fig. 47 is substantially the same as that of the embodiment shown in Fig. 6, except that the pump control valve 1076 is pressure balanced when in the closed position and blocks the flow of fuel between the fuel supply and the pump chamber, thus allowing the control scheme flexibility discussed with respect to the embodiment disclosed in Figs. 44-45.

Referring now to Figures 48-51, another embodiment of the present invention is disclosed. Referring to Figure 48, pump control valves 1100 and 1102 are vertically mounted in respective valve recesses 1104 and 1106 formed in the upper surface 1108 of accumulator housing 1110. Pump control valves 1100 and 1102 are preferably each a solenoid-operated valve assembly of the type disclosed in commonly assigned U.S. Patent No. 4,905,960 to Barnhart. Pump units 1112 and 1114 are mounted in respective pump unit recesses 1116 and 1118 formed in the lower surface of accumulator housing 1110 directly below respective valve recesses 1104 and 1106. The fuel passages in the accumulator housing 1110 associated with each pump control valve 1100 and 1102 are of similar construction and therefore only one group of passages and components will be described below.

Referring to Fig. 49, a pump outlet passage 1120 extends from the valve recess 1104 to the pumping chamber of the pump unit 1112 to form a valve cavity for receiving a valve element 1122 of the pump control valve 1100. A drain passage 1124 extends transversely inwardly from one side of the accumulator housing 1110 into communication with the pump outlet passage 1120. The open end of the drain passage 1124 is fluid-tightly closed with a plug 1126. A pump unit check valve 1128 is disposed in the drain outlet 1124 and is adapted for sealing engagement with an annular valve seat surrounding the drain passage 1124. A vertical passage 1132 extends from the lower surface of the accumulator housing 1110 through the drain passage 1124 upwardly into communication with a storage chamber 1134d formed in the accumulator housing 1110. A similar vertical passage 1133 associated with the pump unit 1114 connects a corresponding discharge passage (not shown) to the storage chamber 1134d. A main supply passage 1136 formed in the pump housing 1138 supplies low pressure fuel to the pump control valve 1100 via a connecting passage 1140 and a side channel 1142. A similar side channel 1143 extends from the connecting passage 1140 for supplying fuel to the other pump control valve 1102. It should be noted that while the pump units 1112 and 1114 are shown to be similar to the embodiment disclosed in Fig. 40 and described above, the pump units may take the form of another embodiment.

Referring now to Figs. 50 and 51, the accumulator housing 1110 of the embodiment illustrated in Figs. 48-49 includes an upper row of elongated accumulator chambers 1134a-d (Fig. 50) and a lower row of elongated accumulator chambers 1134e-g. Each of the accumulator chambers is formed by longitudinal bores through the accumulator housing 1110 from an end wall 1144. The open end of each accumulator chamber is fluid-tightly sealed by the corresponding plug 1146. The upper row of accumulator chambers is connected by a first transverse passage 1148 which extends transversely from one side of the accumulator housing 1110 through each of the accumulator chambers 1134a-d. The accumulator housing 1110 further includes a pair of recess drain passages 1150 and 1152 extending from respective pump unit recesses 1116 and 1118 for directing fuel leakage collecting in the respective recess spaces 1154 and 1156 to a main drain passage 1158. As shown in Fig. 50, the accumulator chamber 1134c terminates approximately midway in the accumulator housing 1110 adjacent the first transverse passage 1148. The accumulator chambers 1134e-g are also separated by a second transverse passage 1160 (Fig. 51) which extends across the accumulator housing 1110 in the same vertical plane as the first transverse passage 1148. The upper and lower rows of accumulator chambers are connected by a vertical passage 1162 which extends from the second transverse passage 1160 into communication with the accumulator chamber 1134c. A fuel supply passage 1164 extending from the lower surface of the accumulator housing 1110 also communicates with the accumulator chamber 1134c. A recess 1166 formed in the open end of the fuel supply passage 1164 is adapted to receive a fuel feed tube 1169 (Fig. 48) which directs the fuel temporarily collected in the storage chambers to the fuel injector(s) (not shown) for delivery to the engine via a manifold (not shown), as previously described in connection with various other embodiments.

Referring now to Figs. 52 and 53a, there is shown another embodiment of the present invention which is the same as the previous embodiment of Figs. 48 and 49, except that an accumulator 1168 is spaced from a pump head 1170. The pump head 1170 does not include any accumulator chambers, but only an elongated common discharge passage 1172 connected to vertical passages 1132, 1133 for receiving high pressure fluid from each pump unit 1112, 1114. The accumulator 1168 includes an accumulator housing 1174 defining a generally cylindrical accumulator chamber 1176. However, the accumulator 1168 may include a plurality of interconnected accumulator chambers similar to the embodiments of Figs. 7 and 50. One end of the storage chamber 1176 is fluid-tightly closed with a plug having a stepped recess 1180 for receiving a pressure sensor 1182. A central passage 1184 connects the stepped recess 1180 to the storage chamber 1176, allowing the pressure sensor 1182 to monitor the fuel pressure in the storage chamber 1176. The opposite end of the storage chamber 1176 is fluid-tightly closed with an intermediate piece 1186 having an internal recess 1188. The intermediate piece 1186 also includes an inlet passage 1190 and an outlet passage 1192 extending from the inner end of the internal recess 1188. A fuel delivery tube 1194 is connected at one end to the common delivery passage 1172 and at an opposite end to the inlet passage 1190 for delivering fuel from the common delivery passage 1172 to the storage chamber 1176. A fuel supply tube 1196 is connected at one end to the outlet passage 1192 for delivering high pressure fuel from the storage chamber 1176 to the injection control valve (not shown). The open ends of the common delivery passage 1172, the inlet passage 1190 and the outlet passage 1192 include respective recesses 1198 with a tube seat 1200 for engagement with a tube head 1202 formed on the end of the respective tube 1194, 1196. Each recess 1198 has internal threads for engagement with a complementary external thread formed on a substantially cylindrical tube fitting 1204. Each tube 1194, 1196 passes through the corresponding tube fitting 1204 so that one end of the tube fitting 1204 rests against the tube head 1202. Rotation of the tube fitting 1204 relative to the recess 1198 and the corresponding tube 1194, 1196 presses the tube head 1202 inwardly into sealing engagement with the tube seat 1200, thereby establishing a fluid-tight connection between the corresponding passages 1172, 1190, 1192 and the corresponding tube 1194, 1196.

The remotely mounted accumulator design of Figures 52 and 53a allows for the accumulator 1168 to be mounted in a potentially more convenient/advantageous location around the engine. Furthermore, the size of the pump head 1170 is reduced both in the longitudinal direction as shown in Figure 52 and in the transverse direction as shown in Figure 53a. This smaller size of the pump head results in a more compact arrangement that can better match the enclosure limitations of certain engine or vehicle designs.

Reference is now made to Fig. 53b, which discloses another embodiment of the present invention which is the same as the previous embodiments of Figs. 52 and 53a, and therefore the same components are provided with the same reference numerals. In this embodiment, a separately formed storage housing 1187 is attached to a pump head 1189. The storage housing 1187 has a substantially cylindrical shape and comprises a storage chamber 1191 with a closed end 1193 and an open end 1195. The open end 1195 is screwed into a recess 1197 formed in an end wall 1199 of the pump head 1189 for a fluid-tight connection between the storage housing 1187 and the pump head 1198. A common delivery passage 1172 extends through the pump head 1198 into communication with the recess 1197 and the accumulator chamber 1191 for delivering high pressure fuel from the pump units 1112 and 1114 to the chamber 1191. The pressure sensor 1182 is located in a recess 1201 formed in the closed end 1193 and connected to the accumulator chamber 1191 by a passage 1203. The arrangement of Fig. 53b is particularly advantageous because it provides a compact modular high pressure fuel pump assembly with an accumulator that is inexpensive to manufacture and easy to mount to the assembly.

Reference is now made to Figures 54a and 54b which disclose gap filter assemblies which serve to capture small foreign particles in the fuel flowing from the accumulator to the injection control valve (not shown). It is known that the meshing gears of a gear pump, such as the booster pumps 406 and 494 shown in Figures 28 and 30 respectively, frequently come into contact with one another as they mesh during normal operation and form small metal particles. If not captured by the booster pump filter, these metal particles are carried by the fuel through the fuel pumping system. However, these particles have been found to interfere with successful operation of the injection control valve and manifold of the present invention. Both the injection control valve and manifold rely on extremely small clearances between their components so that one or more of the components can be moved relative to another while creating a fluid-tight seal at the clearance. Foreign matter in the fuel gets stuck between the components in these gaps and causes excessive abrasion or even blockage of the moving part and possibly the gradual loss of Fluid seal. Therefore, it is desirable to arrange a filter in the fuel path upstream of the injection control valve, which is capable of removing small particles from the fuel.

Fig. 54a discloses a rod filter assembly 1206 disposed along the fuel flow path between the accumulator 1208 and the injection control valve (not shown). The rod filter assembly 1206 includes a rod filter 1210 disposed in a filter cavity 1212 formed in one end of a fuel feed tube 1214 of a feed tube mounting assembly 1216. The pipe attachment arrangement 1216 is the same as the pipe fitting connections previously described in connection with the embodiments of Figs. 5 and 52, except that the end of the supply pipe 1214 includes the filter cavity 1212 which is large enough to accommodate the rod or slot filter 1210. As shown in Fig. 54b, the rod or slot filter may be located in a filter housing 1218 disposed along a fuel supply pipe 1220. In this case, conventional high pressure pipe attachment arrangements 1222 are used to attach each end of the supply pipe 1220 to a corresponding end of the filter housing 1281. In both embodiments of Figures 54a and 54b, the rod or edge filter 1210 serves to advantageously prevent small particles from flowing through the fuel system downstream of the accumulator 1208, thereby preventing foreign objects from causing abrasion and/or damage to the injection control valve and distributor.

Reference is now made to Figs. 55a-55c, which disclose various other embodiments of the accumulator of the present invention. The accumulators previously discussed with reference to the preceding embodiments of the present invention all include an accumulator housing having an accumulator chamber with an open end that is fluid-tightly closed by a plug having an external thread for engagement with a complementary internal thread formed on the inner surface of a recess formed in the open end of one or more chambers. Although such threaded connections also include some type of seal, such as an O-ring, such Sealed screw connections may develop a leak at particularly high fuel pressures, allowing fuel to escape from the storage chamber, resulting in an undesirable loss of fuel pressure in the storage tank, which adversely affects the metering of the fuel.

Figures 55a-55c disclose alternative embodiments of the accumulator that prevent fuel leakage from the ends of the accumulator chambers. Figure 55a discloses an accumulator housing 1230 that includes a stepped recess 1232 formed in one end of the housing 1230. The accumulator chambers 1234 are formed by a bore through an inner end wall 1236 of the stepped recess 1232. An end plate 1238 is then disposed in the stepped recess 1232 against a step 1233 formed by the stepped recess 1232. The end plate 1238 can then be securely and sealingly connected to the accumulator housing 1230 by welding along a circumferential seam 1240 formed between the outer circumferential edge of the end plate 1238 and the edge of the accumulator housing 1230 defining the open end of the stepped recess 1232. A common flow cavity is formed between the inner end wall 1236 and the inner surface of the end plate 1238 so that fuel flow between the accumulator chambers 1234 is possible. The welded circumferential seam 1240 is particularly effective in sealing the accumulator chambers 1234. Consequently, this embodiment provides an accumulator housing 1230 with a single welded end plate 1238 that is particularly resistant to fuel leakage.

Fig. 55b discloses another embodiment of the accumulator of the present invention, which is the same as the embodiment disclosed in Fig. 55a, except that a second stepped recess 1242 is formed at the opposite end of the accumulator housing 1230 for receiving a second end plate 1243.

Fig. 55c discloses a third embodiment of the accumulator of the present invention, which comprises an accumulator housing 1244 formed by the welded connection of a first accumulator block 1246 and a second accumulator block 1248. The accumulator chambers and any other longitudinal passages are formed in each block 1246, 1248 by the corresponding end walls 1250, 1252 before blocks 1246, 1248 are joined. The end walls 1250, 1252 are then placed in place to form a peripheral joint 1254 that runs around the entire accumulator housing. The peripheral joint is then welded so that blocks 1246 and 1248 are securely fastened together while forming a seal that prevents fuel leakage from the accumulator chamber (not shown). The embodiments shown in Figs. 55a-55c significantly reduce the likelihood of fuel leakage from those areas of the accumulator housing used to form the accumulator chambers.

Reference is now made to Figures 56-62, which disclose several devices that can be incorporated into the fuel system of the present invention to provide a fuel flow control function. By reducing the rate at which fuel pressure at the nozzle assembly increases in the initial phase of injection, and thus reducing the initial amount of fuel injected into the combustion chamber, the various embodiments of the present invention are better able to achieve various objectives, such as more efficient and complete fuel combustion with lower emissions. The fuel flow control devices discussed below are intended to better enable the present fuel system to meet the ever-increasing demands for reduced emissions.

Referring first to the embodiment illustrated in Fig. 56, a flow control device, generally designated 1260, is disposed along the fuel delivery circuit 1262 between the fuel injection control valve 20 and the manifold 16 of Fig. 1. However, the flow control device 1260 could be used in any of the embodiments of the present fuel delivery system previously disclosed. Likewise, the flow control device 1260 is illustratively disposed in a manifold housing 1264 in Fig. 56. However, the device 1260 may be installed in the fuel delivery circuit 1262 anywhere between the injection control valve 20 and the manifold 16.

As shown in Fig. 56, the quantity control device 1260 includes a flow limiting valve 1266 disposed in the fuel delivery circuit 1262 and a quantity control bypass valve 1268 disposed in a side passage 1270. The flow limiting valve 1266 includes a slidable piston 1272 disposed for sliding movement in a piston chamber 1274 formed in the fuel delivery circuit 1262 to form a fuel inlet 1276 and a fuel outlet 1278. The slidable piston 1272 includes a first end 1280 disposed adjacent to the fuel inlet 1276, a second end 1282 disposed adjacent to the fuel outlet 1278, and a central bore 1284 extending inwardly from the first end 1280 and terminating at an inner end 1286. The slidable piston 1272 also includes a cylindrical outer surface 1288 that can have a sufficiently close sliding fit with the inner surface of the piston chamber 1274 to form a fluid seal between the surface 1288 and the inner surface of the piston chamber 1274. The second end 1282 of the sliding piston 1272 includes a tapered surface 1290 for engagement with an annular valve seat 1292 formed on the distributor housing 1264 at the fuel outlet 1278 when the sliding piston 1272 is moved to the right as shown in Fig. 56.

The slidable piston 1272 also includes a central opening 1294 extending through the second end 1282 to fluidly communicate the central bore 1284 with the fluid outlet 1018 regardless of the position of the slidable piston 1272. A plurality of first phase openings 1296 extend from the central bore 1284 through the second end 1282. The first phase openings 1296 are aligned with respect to the valve seat 1292 such that when the flow restriction valve 1266 is in the position shown in Figure 56, hereinafter referred to as the second phase position, the flow of fuel from the first phase openings 1296 to the fuel outlet 1278 is blocked by the abutment of the conical surface 1290 on the valve seat 1292. The flow limiting valve 1266 includes a spring cavity 1298 between the piston 1272 and the distributor housing 1264 for receiving a tension spring 1300. An annular step 1302 formed on the piston 1272 serves to provide a spring seat for the spring 1300 which biases the piston 1272 to the left as shown in Fig. 56 to a first phase position.

The side passage 1270 communicates at one end with the fuel inlet 1276 through the piston chamber 1274 and at an opposite end with the fuel outlet 1278. The slidable piston 1272 includes radial grooves 1304 in the end face of the first end 1280 that allow fuel flow between the fuel inlet 1276 and the side passage 1270 when the flow limiting valve 1266 is in the first phase position. The flow control bypass valve 1268 is disposed along the side passage 1270 in a flow control valve cavity 1306. The flow control recirculation valve 1268 includes an elongated valve element 1308 having a conical valve surface 1310 for engagement with an annular valve seat 1312 formed in the manifold housing 1264. The flow control recirculation valve 1268 is preferably a pressure balanced, solenoid actuated, two-position, two-way valve that includes a bias spring 1314 arranged to bias the valve element 1308 to the closed position against the valve seat 1312. A solenoid assembly, shown at 1316, is used to move the valve element 1308 to the right in Fig. 56 to a full flow open position with the conical valve surface 1310 separated from the annular valve seat 1312, thereby allowing flow through the side channel 1270. The quantity control circulation valve 1268 can alternatively be hydraulically operated.

Generally, the flow limiting valve 1266 operates to control or regulate the pressure rate increase at the nozzle assembly during the initial phases of an injection event, as illustrated by phases I and II in Fig. 57, and also controls the return flow of fuel through the delivery circuit at the end of the injection event when the injection control valve 20 is connected to the outlet, thereby minimizing cavitation in the fuel delivery circuit and associated fuel injection lines. The quantity control bypass valve 1268 operates primarily to ensure a rapid increase in the pressure rate. possible when it is desired to achieve maximum pressure at the nozzle assembly by providing an unrestricted flow path through the fuel delivery circuit 1262 after the initial injection period, as illustrated by Phase III in Fig. 57.

In particular, during operation, immediately prior to the start of an injection event, the injection control valve 19 is in the closed position in which the fuel delivery circuit 1262 is connected to the outlet. At this time, the flow restriction valve 1266 is in its first phase position with the first end 1280 against the distributor housing 1264 and allowing fluid communication between the fuel inlet 1276 and the fuel outlet 1278 via both the central opening 1294 and the first phase openings 1296. The flow control bypass valve 1268 is in the closed position under the force of the tension spring 1314 in which flow through the side channel 1270 is blocked. Once the injection control valve 20 is energized to connect the accumulator pressure to the fuel delivery circuit 1262, high pressure fuel initially flows through both the center orifice 1294 and the first phase orifices 1296, creating an initial pressure rise downstream of the flow restriction valve 1266 and at the corresponding nozzle assembly, as shown by Phase I in Fig. 57. However, the accumulator fuel pressure at the fuel inlet 1276 acts on the end face of the first end 1280 and on the inner end 1286 of the center bore 1284 to move the sliding piston 1272 to the right in Fig. 56, thereby moving the sliding piston 1272 to the second phase position in which the conical surface 1290 is in contact with the valve seat 1292. Thus, fuel flow through the first phase orifices 1296 is blocked while a limited amount of fuel passes through the middle orifice 1294 to the fuel outlet 1278, thereby reducing the rate at which fuel pressure at the nozzle assembly increases, as illustrated by Phase II in Figure 57. After a predetermined period of time, and preferably prior to the middle portion of the injection event, the quantity control bypass valve 1268 is switched to the open position, allowing full fuel flow through the side passage 1270, thereby causing a sharp increase in fuel delivery pressure, as illustrated by the upward increasing pressure rate of Phase II is shown in Fig. 57. The pressure at the nozzle assembly quickly reaches a maximum value by the end of the injection event, which is determined by the closure of the injection control valve 20. Consequently, as shown in Fig. 57, the quantity regulating device 1260 produces a first phase of fuel injection (Phase I) with a high pressure rate increase, a second phase of fuel injection (Phase II) with a reduced pressure rate which is less than in Phase I, and a third phase in which the pressure rate increase is initially greater than in Phase II. By reducing the pressure rate increase at the nozzle assembly during the initial injection phases, i.e. Phase II, the quantity regulating device 1260 also reduces the amount of fuel delivered to the combustion chamber during the initial phase, which in turn advantageously reduces the level of emissions produced by the combustion event.

Once closed, the injection control valve 20 blocks the fuel from the accumulator while connecting the fuel delivery circuit 1262 to the outlet. After a predetermined period of time, the quantity control bypass valve 1258 is de-energized and moved to the closed position by the tension spring 1314. It should be noted, however, that the pressure relief of the fuel delivery circuit 1262 downstream of the quantity control device 1260 can be controlled or regulated in a variety of ways depending on the timing of the closure of the quantity control bypass valve 1268 relative to the closure of the injection control valve 20. If the closure of the quantity control bypass valve 1268 is delayed or retarded by a substantial amount of time after the closure of the fuel injection control valve 20, the side channel 1270 serves as the primary relief channel, allowing a strong backflow of fuel to the outlet, thus quickly relieving a substantial amount of fluid pressure from the downstream delivery circuit and the corresponding fuel injection line, while a secondary relief flow is established through the flow restriction valve 1266. However, by closing the quantity control bypass valve 1268 at the same time as or immediately after closing the injection control valve 20, the primary relief is carried out by the flow limiting valve 1266. In both cases, the fuel pressure at the fuel inlet vent 1276 is less than the fuel pressure in the fuel outlet 1278 as soon as the quantity control bypass valve 1268 closes. As a result, the fluid forces acting on the end face of the piston 1272 at the second end 1282, in combination with the biasing force of the spring 1300, become greater than the fluid forces acting on the piston 1272, which tend to move the piston 1272 to the right in Fig. 56. Consequently, the slidable piston 1272 of the flow limiting valve 1266 immediately moves to the left in Fig. 56 to the first phase position in which the first phase ports 1296 are connected to the fuel outlet 1278, allowing fuel flow through the flow limiting valve 1266 through the ports 1294 and 1296. The center orifice 1294 and the first phase orifices 1296 are of sufficiently large diameter such that their combined cross-sectional flow area creates the necessary backflow during the run-off process to provide sufficient fuel pressure relief at the nozzle assembly to avoid secondary injections. On the other hand, the center orifice 1294 and the first phase orifices 1296 are small enough to provide a combined flow area designed to limit backflow to a predetermined value necessary to minimize cavitation in the circuit and injection lines between the flow limiting valve 1266 and the nozzle assemblies. Therefore, the flow restriction valve 1266 serves as a variable flow valve when it is moved between the first phase and second phase positions, advantageously utilizing the flow restriction feature of the center orifice 1294 during the injection event to regulate the pressure rate increase, while advantageously controlling the backflow during the run-off event to both prevent secondary injections and minimize cavitation.

It should be noted that a single fixed orifice located in the main stream causes a significant injection delay. Much of this delay is eliminated by the present quantity control device, which includes a central orifice 1294 in a movable piston 1272. The displacement of this piston leads to practically no differential in of the pressure record compared to an unloaded line. This value depends largely on the displacement of the plunger and the volume of the pressurized system. If the geometry ("d" diameter and "s"stroke; Fig. 56) of the plunger 1272 is properly dimensioned, the pressure can be maintained slightly less than the opening pressure of the injector. This means that the invisible part of the injection rate has a "fast response time" (no delay) and the orifice 1294 begins to dominate the process at exactly this pressure value to regulate the rate.

A further advantage of this design is achieved by locating the quantity control bypass valve 1268 downstream of the injection control valve. This arrangement minimizes the leakage loss that occurs through the valve 1268. This leakage loss is four times less than if the valve 1268 were located upstream of the injection control valve (assuming that the duration is 30 degrees crank angle and the engine is a six-cylinder four-stroke engine).

Referring now to Figures 58 and 59, another quantity control device 1320 is disclosed in connection with the present fuel pumping system of the present invention which includes a high pressure accumulator 12, an injection control valve 20 and a manifold 16 positioned along the fuel delivery circuit 1322 for delivering precise amounts of fuel through injection lines 1324 for delivery to the engine cylinders (not shown) via respective nozzle assemblies 11. The quantity control device 1320 includes a high pressure delivery passage 1328 of the fuel delivery circuit 1322 which connects the accumulator 12 to the injection control valve 20. At the beginning of the injection process, when the injection control valve 20 moves to an open position in which the accumulator 12 and the high pressure discharge passage 1328 are connected to the fuel feed circuit 1322 downstream of the injection control valve 20, there is an immediate drop in the fuel pressure in the high pressure discharge passage 1328 immediately upstream of the injection control valve 20, while a high pressure fuel pulse from the accumulator 12 travels rapidly from the accumulator to this low pressure region and then on to the nozzle assembly 11. Therefore, there is a time delay between the opening of the injection control valve 20 and the arrival of the high pressure pulse at the injection control valve 20. The greater the distance that the fuel pulse must travel from the accumulator 12 to the injection control valve 20, the longer the time required to increase the fuel pressure at the control valve, and thus in the fuel injection line adjacent to the nozzle assembly, to the pressure rate necessary to achieve the optimum high fuel pressure. Therefore, by increasing the distance between the accumulator 12 and the injection control valve 20, i.e., by increasing the high pressure delivery passage 1328, the quantity control device 1320 of the present invention slows the pressure rate increase at the nozzle assembly, as shown by the pressure-time curve in Fig. 59.

Referring now to Fig. 60, another flow control device 1330 is disclosed which is similar to the embodiment shown in Fig. 58 in that a high pressure delivery loop 1332 of a length is used to control the time required for full, unrestricted accumulator flow to result in the high pressure reaching the nozzle assembly 11. In this embodiment, however, an orifice 1334 is disposed in a restricted flow passage 1336 so that the high pressure delivery loop 1332 serves as a bypass around the restricted flow passage 1336. Again, as in the previous embodiment, the flow control device 1330 takes advantage of the fact that pressure waves take time to propagate through the high pressure delivery loop 1332, thereby delaying the arrival of the high pressure at the nozzle assembly 11 and providing an initial period of injection with a low pressure rise rate. In addition, however, the orifice 1334 serves to slow the rate of pressurization at the nozzle assembly to the desired pressure rate. Therefore, the orifice 1334 can be selected with a predetermined cross-sectional flow area that provides a desired pressure rate during the initial injection period. Furthermore, the orifice 1334 serves to dampen the undesirable pressure waves that fluctuate in the lines between the accumulator and the injection control valve. Referring to Figure 59, although for a given length of the high pressure delivery loop 1332 the time delay (T) would remain constant, the pressure rate could be adjusted by selecting a properly sized opening 1334 to produce a desired pressure rate change, as shown by the dashed lines 1338.

Reference is now made to Fig. 61, which discloses a flow control device 1340 which is the same as the flow control device 1330 of Fig. 60, except that a flow control or flow control valve 1342 is disposed in a high pressure side passage 1344 to direct flow around the orifice 1334. Preferably, the flow control valve 1342 is a pressure balanced, solenoid-operated, two-position, two-way valve which can be disposed in a closed position in which flow through the high pressure side passage 1344 is blocked and in an open position in which flow is possible. The quantity control valve 1342 enables precise control and modification of the time delay shown in Fig. 59 by electronically controlling and adjusting the opening and closing of the quantity control valve 1342.

The flow control devices shown in Figures 56-62 and discussed previously can be connected to nozzle assemblies, such as the two-spring nozzle assembly manufactured by Bosch or the plunger in the nozzle assembly as developed by AVL, which are designed to reduce the amount of fuel delivered during the first part of the injection. When these nozzle assembly designs are connected to the accumulator flow control concepts of the present invention, the combination of the two results in further reductions in the amount of fuel injected at the beginning of the injection event.

Reference is now made to Figs. 62a and 62b, which disclose a quantity control coupling 1350 for incorporating the quantity control devices disclosed in Figs. 60 and 61 into a fuel system, which at the same time represents a housing for receiving a rod or gap filter. The quantity control coupling 1350 comprises a substantially cylindrical housing 1352 with an inlet part 1354, an outlet circulation part 1356 and a central supply bore 1358, which extends both through the inlet portion 1354 and the outlet recirculation portion 1356. The housing 1352 further includes a recirculation return portion 1360 and a discharge portion 1362 integrally formed with the inlet portion 1354 and the outlet recirculation portion 1356. The discharge portion 1362 includes a supply passage 1364 extending through the portion 1362 inwardly to the central supply bore 1358. A flow restriction opening 1366, similar to opening 1334 of Figs. 60 and 61, is disposed at the inner end of the supply passage 1364 to connect the supply passage 1364 to the central supply bore 1358. As shown in Fig. 62b, the recirculation return portion 1360 includes a return passage 1368 extending through the housing 1352 and connected to the supply passage 1364 downstream of the opening 1366. Referring again to Figs. 62a and 62b, the inlet portion 1354 is connected by a high pressure tube fitting 1370 to a fuel supply tube 1372 which delivers fuel from the reservoir (not shown). The outlet recirculation portion 1356 is connected to one end of a recirculation loop or tube shown at 1374, while the opposite end of the recirculation loop 1374 is attached to the recirculation return portion 1360. The recirculation loop 1374 corresponds to the delivery loop 1332 and side channel 1344 disclosed in Figs. 60 and 61, respectively. Therefore, the flow control valve 1342 of Fig. 61 may be disposed along the circulation loop 1374. Also, a rod filter 1376 is disposed in the central supply bore 1358 of the housing 1352 adjacent the inlet portion 1354. A retaining pin 1377 is disposed in the central bore 1358 in interference fit between the rod filter 1376 and one end of the supply tube 1372 to secure the rod filter 1376 in the central supply bore 1358. The retaining pin 1377 includes axial grooves 1379 which permit fuel flow through the central supply bore 1358 to the circulation loop 1374. The rod or gap filter 1376 serves to remove small particles such as metal shavings from the fuel to prevent the particles from reaching the injection control valve and manifold located downstream. Therefore, the flow control coupling 1350 provides a compact, effective means for implementing the flow control devices of Figs. 60 and 61 while simultaneously providing an easily accessible, yet effective housing for a rod or gap filter.

Reference is now made to Figures 63a-69, which disclose various devices for minimizing cavitation in the fuel delivery circuit and high pressure injection lines while minimizing the possibility of secondary injection. The cavitation, i.e., vapor pockets or bubbles, in the delivery circuit and injection lines leading to the nozzle assemblies causes insufficient injection pressure and unpredictable, uncontrollable variations in both fuel quantity and injection timing. Cavitation is particularly common in high pressure lines of fuel systems where such lines are connected cycle after cycle to a low pressure outlet, as in the fuel pumping system of the present invention. The following devices advantageously control cavitation by 1) minimizing cavitation by limiting fuel back or reverse flow during the run-off event and/or 2) filling the injection lines with fuel after each run-off event and before the following injection event. In particular, the cavitation control devices illustrated in the embodiments of Figures 64a-64e minimize cavitation by limiting fuel back flow during the run-off event, while the devices disclosed in Figures 63a, 63b and 69 minimize the effects of cavitation primarily by filling the downstream lines with fuel.

Referring initially to the embodiment disclosed in Figures 63a and 63b, a cavitation control device, generally designated 1400, is formed in a manifold housing 1402 of a manifold 1404. Figure 63a also shows an injection control valve 1406, a low pressure accumulator 1408 mounted in a spacer housing 1410, a two-piece gear pump housing 1412, 1414, and a pressure booster or gear pump 1416. These various components are essentially the same as in the embodiment previously described with reference to Figure 30, except for the addition of a cavitation control device 1400. The cavitation control device 1400 includes an axial passage 1418 leading from the outlet of the pressure booster pump 1416 adjacent to the low accumulator 1408 passes through the spacer housing 1410, the two-piece gear pump housing 1412, 1414 and the distributor housing 1402. The axial passage 1418 terminates approximately in the middle of the distributor housing 1402 for communication with a discharge passage 1420 which extends radially inward at an angle through the distributor housing 1402 and a fixed shaft sleeve 1422 which surrounds a distributor rotary shaft 1424. The innermost end of the discharge passage 1420 is in continuous communication with an annular groove 1426 formed in the outer surface of the distributor shaft 1424. A transverse passage 1428 extends diagonally from the annular groove 1426 through the central axis of the distributor shaft 1424 to the opposite side of the distributor shaft 1424. The transverse passage 1428 connects the annular groove 1426 to a refill opening 1430 formed in the outer surface of the distributor shaft 1424. As shown in Figs. 63a and 63b, the refill opening 1430 is disposed in a common vertical plane with an injection port or window 1432 which sequentially communicates with fuel receiving passages 1434 evenly spaced about the circumference of the rotor bore 1436. As previously discussed with reference to the embodiment of FIG. 5, during the opportunity phase, the injection control valve 1406 delivers fuel through a fuel delivery circuit to the injection port 1432 to create an injection event. The fuel delivery circuit includes passages 1438 and 1440 formed in the distributor housing 1402 and the shaft sleeve 1422, respectively, an annular supply groove 1442 formed in the distributor shaft 1424, and a delivery passage 1444 extending from the annular supply groove 1442 diagonally through the distributor shaft 1424 into communication with the injection port 1432. As shown in Fig. 63b, at the end of the injection event, as the distributor shaft 1424 rotates clockwise, the injection port 1432 moves out of communication with a particular fuel receiving passage 1434. As the distributor shaft 1424 continues to rotate, the refill port 1430 is moved into fluid communication with the receiving passage 1434 through which an injection event previously occurred. This directs low pressure fuel from the outlet of the pressure booster pump 1416 to the corresponding fuel receiving passage 1434 via the passages 1418, 1420, the annular groove 1426 and the transverse passage 1428. Each fuel receiving passage 1434 is connected to a nozzle assembly 1445 of an associated engine cylinder through a corresponding injection passage 1446 formed in the manifold housing 1402, a corresponding injection bore 1448 formed in an outlet fitting 1450, and a corresponding injection line 1452 connected at one end to the outlet fitting 1450 and at an opposite end to the nozzle assembly 1445. In this manner, the cavitation control device 1400 ensures that each injection circuit connecting the manifold 1404 to a corresponding nozzle assembly is replenished with low pressure fuel prior to the next injection event, thereby minimizing variations in fuel quantity and injection timing due to cavitation. Furthermore, since the boost pump fuel pressure is maintained at a relatively constant value, all injection lines are pressurized to approximately the same fuel pressure value for each injection event, thereby improving the predictability of fuel metering and timing.

63a and 64a also show another device for minimizing cavitation, generally designated A. This embodiment includes a backflow limiting valve 1460 disposed along the fuel delivery circuit 1462 between the injection control valve 1406 and the manifold 1404. The backflow limiting valve 1460 includes a movable valve element 1464, an insert 1466, and a retaining ring 1468 received in a recess 1470 formed in the manifold housing 1402. The inner end of the recess 1470 communicates with one end of the passage 1438 through an outlet 1463 for delivering fuel to the manifold 1404. A delivery passage 1472 formed in an injection control valve housing 1474 includes an inlet 1475 arranged to open into the recess 1470 when the injection control valve housing 1474 is arranged adjacent the distributor housing 1402. A spacer plate 1476 is arranged between the injection control valve housing 1474 and the distributor housing 1402. The spacer plate 1476 includes an opening 1478 through which the backflow limiting valve 1460 extends. The retaining ring 1468 is mounted against the inner end of the recess 1470 around the outlet 1493 to support the inlet. insert 1466. The insert 1466 is disposed in the recess 1470 in press fit with the retaining ring 1468 at one end and the injection control valve housing 1474 at an opposite end. The insert 1466 includes an annular base 1480 disposed in fit with the retaining ring 1468 and wall portions 1482 extending upwardly from the base 1480 into fit with the housing 1474. The wall portions 1482 form a valve cavity 1484 for receiving the valve element 1464. A bore 1486 extending through the base 1480 connects the outlet 1463 to the valve cavity 1484. Radial grooves 1488 formed in the upper portion of the base 1480 extend radially outward from the bore 1486 into communication with corresponding slots 1490 separating the wall portions 1482.

The movable valve element 1464 is substantially tubular in shape and is provided with an appropriate outer diameter to allow movement in the valve cavity 1484 along a vertical axis, while the wall portions 1482 provide lateral support for the valve element 1464. A valve seat 1492 formed about the inlet opening is designed for sealing engagement with the valve element 1464 when the valve element 1464 is moved upwardly to a limiting position. The valve element 1464 can be moved downwardly into contact with the inner surface of the cavity 1484 to an open position, as shown in Figure 64. The valve element 1464 is also provided with a suitable width to form an axial gap 1493 to allow fuel flow from the inlet 1475 to the slots 1490 when the valve element 1464 is in the open position. The valve element 1464 includes a central opening 1494 for establishing fluid communication between the inlet 1475 and the outlet 1463 when the valve element 1464 is in the restricting position.

The high pressure connections formed by the seating of the injection control valve housing 1474, the spacer plate 1476 and the distributor housing 1402 are sealed using several devices to prevent high pressure fuel leakage. First, an annular sealing ring, i.e. a C-ring, 1496 is press-fitted between the injection control housing 1474 and the distributor housing 1402 in the opening 1478. In addition, opposing annular fuel collection grooves 1498 are formed in each housing 1474, 1402 radially outward of the seal ring 1496 for collecting fuel that leaks through the seal ring 1496. A drain passage 1500 extends from a fuel collection groove to drain the collected fuel to the drain (not shown). A balance passage 1502 extends through the spacer plate 1476 to connect the opposing fuel collection grooves 1498, allowing fuel collected in both grooves to be directed to the drain. Third, two opposing annular O-ring grooves 1504 are formed in the housings 1474 and 1402 radially outward of the fuel collection grooves 1498 for additional sealing.

During operation, at the beginning of an injection event, when the injection control valve 1406 moves to an open position in which high pressure fuel is directed from the accumulator (not shown) to the delivery passage 1472, the valve element 1464 of the backflow limiting valve 1460 moves under the force of the high pressure fuel into abutment against the inner surface of the valve cavity 1484 to a full flow open position. In this open position, fuel from the delivery passage 1472 flows through the axial gap 1493, the slots 1490 and into the bore 1486 for delivery to the manifold 1404 via the outlet 1463 and the passage 1438. The fuel from the delivery passage 1472 also flows through the central opening 1494 for delivery to the manifold. The valve element 1464 is sized so that the effective flow area of the axial gap 1493 in conjunction with the effective flow area of the central opening 1494 provides a substantially unlimited flow through the restriction valve 1460. At the end of the injection event, when the injection control valve 1406 moves to a run-out position in which the delivery passage 1472 is connected to the run-out, the fuel pressure in the delivery passage 1472 immediately becomes less than the pressure in the passage 1438 and bore 1486. As a result, a reverse fuel flow from the passage 1438 and other downstream passages including the corresponding fuel injection line flows in the reverse direction through the flow restriction valve 1460 to the injection control valve 1406. As previously discussed, without the Use of the flow restriction valve 1460 can cause vapor pockets or bubbles (cavitation) to form in the delivery passages and injection line between the injection control valve 1406 and the nozzle assemblies. However, the reverse flow restriction valve 1460 helps minimize cavitation by allowing movement of the valve element 1464 to a restricting position against the valve seat 1492. In the restricting position, the valve element 1464 blocks reverse fuel flow through the annular gap 1493 while allowing limited fuel flow through the central opening 1494. The central opening 1494 has an effective cross-sectional flow area that allows sufficient backflow of fuel to allow adequate depressurization of the passages between the restriction valve 1460 and the nozzle assembly to allow the nozzle valve element (not shown) of the nozzle assembly to close, resulting in predictable injection timing and metering while simultaneously restricting fuel flow to create optimal back pressure to minimize cavitation.

Referring now to Fig. 64b, another embodiment of the flow restriction valve is disclosed which is similar to the embodiment of Fig. 64a in that the valve element 1464 is disposed with the central opening 1494 in a recess 1470 formed in the manifold housing 1402. However, in the embodiment shown in Fig. 64b, the wall portions 1510 are formed integrally with the manifold housing 1402 in the inner end of the recess 1470. The wall portions 1510 extend radially inward to define a central bore 1512 which is connected to the outlet passage 1514 for directing fuel to the manifold 1404. The wall portions 1510 are separated by slots 1516 which communicate with the central bore 1512. In this embodiment, the valve element 1464 is dimensioned to form both an axial gap 1518 between its upper flat surface and the annular valve seat 1492 and an annular radial gap 1520 between its outer peripheral surface and the inner surface of the recess 1470. When in the full flow open position, as shown in Fig. 64b, fuel flows from the delivery passage 1472 through the axial gap 1518 and the radial gap 1520 via the slots 1516 into the central bore 1512 for delivery to the manifold 1404 via the outlet passage 1514. The valve element 1464 functions in the same manner as described with reference to the embodiment of Fig. 64a when it is moved to a restricting position against the annular valve seat 1492 to restrict fuel backflow, thereby slowing the pressure drop in the fuel delivery circuit and injection lines between the valve element 1464 and the nozzle assembly, thereby avoiding excessive cavitation. It should also be noted that this embodiment does not include a spacer plate 1476. Further, the sealing ring 1496 is disposed in a single annular groove 1522 formed in the injection control valve housing 1474. Likewise, only a single fuel collection groove 1524 and a single O-ring groove 1526 are required to receive the O-ring 1528 since only one high pressure connection is formed between the housings 1474 and 1402.

Reference is now made to Fig. 64c, which shows another embodiment of the cavitation control device, which is the same as the embodiment shown in Fig. 64b, except that a conical shaped recess 1530 is formed on the upstream side of a movable valve member 1532 adjacent to the annular valve seat 1492. The central opening 1534 extends through the movable valve member 1532 and connects the conical shaped recess 1530 to the central bore 1512. The conical shaped recess 1530 serves to reduce the surface area of the valve member 1532 that contacts the valve seat 1492, thereby improving the seating of the valve member 1532 against the valve seat 1492.

Referring to Fig. 64d, a fourth embodiment of the backflow limiting valve is disclosed which includes a cylindrical connecting tube 1540 disposed in a recess 1542 formed in both the distributor housing 1402 and the injection control valve housing 1474. The connecting tube 1540 is preferably fixedly secured to the inner wall of the recess 1542 by a press-fit connection, the outer diameter of the connecting tube 1540 being slightly larger than the inner diameter of the portion of the recess 1542 prior to assembly. formed in the manifold housing 1402. The portion of the recess 1542 formed in the injection control valve housing 1474 has an inner diameter slightly larger than the outer diameter of the connecting tube 1540 to provide a clearance therebetween through which fuel leakage can drain to the outlet. The connecting tube 1540 is located at the upstream end of the recess 1542, extends into the manifold housing 1402, and terminates prior to the opposite end of the recess 1542 to form a valve cavity 1544 for receiving a movable valve element 1546. The connecting tube 1540 includes a central bore 1548 to allow fluid flow between the delivery passage 1472 and the valve cavity 1544. The connecting tube 1540 also includes a valve seat 1550 formed on its end wall adjacent the valve cavity 1544 for engagement with the movable valve member 1546. The movable valve member 1546 has a conically shaped recess 1522 formed in one end adjacent the valve seat 1550 and a central opening 1554 extending from the conically shaped recess 1522 through the valve member 1546 into communication with the outlet passage 1556. Inner annular wall portions 1558 formed around the outlet passage 1556 extend to the movable valve member 1546. The wall portions 1558 are separated by slots 1560 extending radially outward from the outlet passage 1556 into communication with an outer annular groove 1562. The axial grooves 1564 are formed in the outer surface of the movable valve element 1546 around its circumference. When the movable valve element 1546 is moved to the open position by the upstream fuel pressure, as shown in Fig. 64d, fuel can flow from the central bore 1548 into the valve cavity 1544 and through the axial grooves 1564 via the annular groove 1562 and the slots 1560 into the outlet passage 1556. The advantages and operation of this embodiment of the backflow limiting valve are the same as the previous embodiment.

Fig. 64e shows another embodiment of the backflow limiting valve of the present invention, which comprises a cylindrical connecting tube 1570 arranged in a recess 1572 similar to that of the previous embodiment. The connecting tube 1570 and a However, retaining ring 1574 is held in an end-to-end press fit within recess 1572. Connecting tube 1570 includes a central bore 1576 communicating at one end with discharge passage 1472 and at an opposite end with discharge passage 1578. In this embodiment, a movable valve member 1580 is disposed within a recess 1582 formed in the upstream end of central bore 1576. Movable valve member 1580 includes a conically shaped recess 1584 formed in its upstream end and a central opening 1586 communicating recess 1584 with central bore 1576. In this embodiment, axial grooves 1588 are formed in the inner surface of connecting tube 1570 along the entire length of tube 1570. Therefore, during the injection event, when the movable valve element 1580 is placed in the full flow open position as shown in Fig. 64e, fuel flows from the passage 1472 through the axial grooves 1588 via the central bore 1576 to the outlet passage 1578. In addition, the movable valve element 1580 is biased to the flow restricting position by a tension spring 1590 disposed in the central bore 1576. The tension spring 1590 assists in moving the valve element 1580 to the flow restricting position upon communication of the fuel delivery passage 1472 with the outlet at the end of the injection event.

Referring now to Figure 65, another embodiment of the cavitation control apparatus of the present invention includes an additional fuel supply, generally designated 1600, delivered to the drain port 1602 of the injection control valve 1604. As previously explained in connection with the fuel system of the present invention, the injection control valve 1604 fluidly connects the accumulator 1606 to the manifold 1608 to define an injection event. The injection control valve 1604 terminates the injection event by connecting the fuel delivery port 1610, and thus the corresponding injection line connected by the manifold 1608, to the drain port 1602, thereby allowing fuel flow from the delivery port 1610 and the injection line 1612 to a drain 1614. As previously noted, this drain event can lead to cavitation in passage 1610 and the corresponding downstream passages. The embodiment shown in Figure 65 minimizes the effects of cavitation in passage 1610 and injection line 1612 during the interruption of the injection event by directing additional fuel at a relatively low pressure, i.e., 300 psi (about 2.068 MPa), to the delivery and injection passages between the injection control valve 1604 and nozzle assembly 1616, thereby refilling the passages prior to the next injection event. The additional fuel also minimizes cavitation by slowing the fuel drain during the run-off event, thereby preventing excessive pressure drop in the downstream passages. In this embodiment, the additional fuel is provided by the pressure booster pump 1618, which directs low pressure fuel to the high pressure pump 1620 for delivery to the accumulator 1606. The additional fuel passage 1622 is connected at one end to the downstream side of the pressure booster pump 1618, for example, directly to the delivery passage 1624, connecting the pressure booster pump 1618 to the high pressure pump 1620. The opposite end of the additional fuel passage 1626 is connected to the outlet passage 1602. A restrictor orifice 1626 is disposed in the outlet passage 1602 downstream of the connection to the additional fuel passage 1622. The restrictor orifice 1626 serves to reduce the amount of additional fuel returned to the outlet 1614, thereby minimizing pumping losses.

Reference is now made to Fig. 66, which shows another embodiment of the cavitation control device of the present invention, which includes a pressure regulator 1630 disposed in the discharge passage 1632 extending from the injection control valve 1634. The pressure regulator 1630 includes a cylinder 1636 defining a cavity 1638 connected at one end to the discharge passage 1632. The pressure regulator 1630 also includes a piston 1640 slidably mounted in the cavity 1638 so that the cavity 1638 is divided into an inlet chamber 1642 for receiving fuel from the discharge passage 1632, and a clamping chamber 1644. The cylindrical outer surface of the piston bens 1640 forms a sufficiently close sliding fit with the inner surface of the cylinder 1636 to form a fluid seal between the surfaces so as to substantially prevent fuel from leaking from the inlet chamber 1642 to the tension chamber 1644. A tension spring 1646 is disposed in the tension chamber 1644 which biases the piston 1640 toward the inlet chamber 1642. A leakage drain passage 1548 is connected to the spring chamber 1644 to direct any fuel that collects in the spring chamber 1644 to the drain. A high pressure relief passage 1650 is connected to the cavity 1638 along the length of the cylinder 1636 between the inlet chamber 1642 and the spring chamber 1644. The tension spring 1646 normally biases the piston 1640 to the left in Fig. 66 so that the cylindrical outer surface of the piston 1640 covers the relief passage 1650 and prevents flow from the discharge passage 1632 to the relief passage 1650 via the inlet chamber 1642. During an injection event, the injection control valve 634 places the accumulator 1652 in fluid communication with the manifold 1654 while blocking the flow of fuel between the fuel delivery circuit 1656 and the discharge passage 1632. During this time, the piston 1640 usually blocks the relief passage 1650 because there is no high pressure fuel in the inlet chamber 1642. Once the injection event is completed and the injection control valve 1634 moves to a run-off position in which the fuel injection passages 1658 and a corresponding fuel injection line 1660 are connected to the run-off passage 1632, high pressure fuel flows through the run-off passage 1632 into the inlet chamber 1642. The high pressure of the fuel in the inlet chamber 1642 acts on the end face 1662 of the piston 1640, creating a force tending to move the piston 1640 to the right in Fig. 66. However, the tension spring 1646 resists the rightward movement of the piston 1640, creating a back pressure in the fuel delivery passages and in the corresponding injection line. Once the pressure of the fuel in the inlet chamber 1642 rises to a predetermined value sufficient to overcome the biasing force of the spring 1646, the piston 1640 moves to the right in Fig. 66 and exposes the high pressure relief passage 1650 so that fuel from the inlet chamber 1642, the delivery passage 1658 and other downstream lines including the injection line 1660 can flow in the opposite direction through the discharge passage 1632 and the relief passage 1650. Once the fuel pressure in the spill passage drops below a predetermined value, the piston 1640 moves to the left in Fig. 66 under the force of the tension spring 1646 and blocks the flow of fuel through the relief passage 1650. The inlet chamber 1642 serves as a reservoir for collecting fuel for refilling the injection lines to minimize the effects of cavitation. The force of the piston 1640 against the accumulated fuel in the inlet chamber 1642 pumps fuel at a predetermined low pressure value into the fuel delivery passages and injection lines, thereby refilling any bubbles or vapor pockets that were unexpectedly formed in the delivery passages and injection lines during the spill operation. Likewise, the effective flow area of end face 1662 and the biasing force of spring 1646 are carefully chosen to produce a run-out action that corresponds to the optimum rate of pressure drop in the injection lines and passages connected to the run-out to minimize cavitation. Similarly, a conventional pressure regulator could be used to maintain back pressure without the benefit of a pooled fuel pre-volume to replenish the injection lines.

In addition, the pressure regulator 1630 of Fig. 66 may be combined with the cavitation control device 1400 of Figs. 63a and 63b to advantageously minimize cavitation. The runoff passage 1632 in Fig. 66, which connects the injection control valve to the pressure regulator 1630, is subject to pressure wave fluctuations due to the repeated release of the relatively high injection pressure into the runoff passage caused by the operation of the injection control valve. These pressure wave fluctuations may be transmitted to the injection lines 1660 during fill, thereby adversely affecting the fill process and subsequent injections. By combining the embodiments of Figs. 63a and 66, the relatively constant fuel pressure of the pressure boost pump 416 of the cavitation control device 1400, which does not have pressure wave fluctuations, is used for more effective filling of the injection lines downstream of the distributor without exposing the injection lines to the to pressure wave fluctuations and the associated adverse effects.

Reference is now made to Fig. 67 which shows another embodiment of the cavitation control apparatus of the present invention which is similar to the previous embodiment and therefore the same components are given the same reference numerals as used in Fig. 66. In this embodiment, a pressure regulator 1666 includes a piston 1668 which is biased by the pressure of fuel supplied from the accumulator 1652 to the inlet chamber 1642. A biasing fluid passage 1670 is connected to the accumulator 1652 at one end and to the biasing chamber 1644 at an opposite end. A biasing pin 1672 is slidably mounted in the biasing fluid passage 1670 adjacent the biasing chamber 1644. An inner end 1674 of the roll pin 1672 extends into the roll chamber 1664 into abutment with one end of the piston 1668. An outer end 1676 of the roll pin 1672 is exposed to the reservoir fuel at extremely high pressure. By selecting the proper effective cross-sectional area of the outer end 1676 of the roll pin 1672, the pressure regulator 1666 can be used in the same manner as the embodiment of Fig. 66 to provide sufficient draining of the fuel delivery circuit and injection lines to terminate injection while both maintaining optimum back pressure necessary to minimize cavitation and directing low pressure fuel to the fuel passage and corresponding injection line during the final part of the draining process to refill the injection passages and lines. In addition, this embodiment includes a refill passage 1678 that connects the drain passage 1632 to each of the fuel injection lines 1660 via the manifold 1654 and refills the injection passages and injection line 1660 between the manifold 1664 and the nozzle assembly after the drain event prior to the next injection event. The refill passage 1678 is connected to each of the injection lines 1600 via passages (not shown) formed in the manifold housing and the rotary shaft, similar to the passages shown in Figs. 63a and 63b with respect to the cavitation control device 1400, except that the discharge passage 1420 is connected to the refill passage 1678. Thus, after an injection event, the refill port 1430, shown in Fig. 63a, connects each injection line in turn to the refill passage 1678 so that fuel in the inlet chamber 1642 can flow to the corresponding injection line. The biased piston 1668 of the pressure regulator 1666 maintains a back pressure in the refill passage 1678 during the injection event when the injection control valve 1634 blocks flow through the drain passage 1632. Thus, the pressure regulator 1666 causes fuel to be pumped back into the fuel injection lines 1660 via the refill passage 1678 to fill any vapor pockets or bubbles that may have formed during the previous interruption of the injection event and prior to the next injection, thereby providing accurate and predictable injection timing. Alternatively, a refill groove 1679 may be formed in the distributor shaft 1424. The refill groove 1679 extends around the circumference of a shaft 1424 for a sufficient angular length to connect the fuel receiving passages 1434 that are not connected to the injection port 1432 during a portion of each injection period. Thus, the refill groove 1679 allows for the refilling of the receiving passages 1434 and the corresponding downstream conduits between injection events and for equalizing the initial fuel pressure in these passages prior to each injection event to provide controllable and predictable fuel metering from one injection period or engine cycle to the next.

Referring to Figure 69, another embodiment of the cavitation control device of the present invention is disclosed. This embodiment combines the spring-loaded pressure regulator 1630 of Figure 66 with the refill passage 1678 disclosed in Figure 67. Therefore, the operation and advantages of this embodiment are substantially the same as the previous two embodiments.

As will be apparent from the discussion previously set forth herein, the present invention advantageously provides a fuel system comprising an electronically controllable high pressure fuel pump assembly having a pump, an accumulator and a distributor connected to an electronic a mechanically actuated pump control valve and an injection control valve attached to the modular assembly to form a highly integrated fuel system which provides better emission control and improved engine performance and which can be designed, built and installed into either an original or existing engine design with minor modifications to the existing designs. This highly integrated fuel system is capable of providing very high injection pressures, i.e. 5000-30000 psi (about 34.47 to 206.8 MPa), and preferably in the range 16000-22000 psi (about 110.3-151.7 MPa), with precise control of the amount and timing of injection depending on various engine conditions, while at the same time providing redundant fail-safe electronic components and improved engine performance at an overall lower cost relative to competing state of the art systems.

The present fuel system also offers the advantage of a highly compact, integrated fuel pump assembly by providing a pump housing having at least one pump cavity oriented in a radial direction and an accumulator secured to the pump housing. Such an accumulator may provide an overhang in either the lateral and/or axial directions and a pump control valve secured to the overhang portion of the accumulator housing adjacent to the pump housing. In addition, the accumulator housing is secured to the pump housing at one end of the pump housing to form a cantilevered lateral overhang such that the overhang provides an offset cross-section for the fuel pump assembly which complements the irregular cross-section of the internal combustion engine to which the fuel assembly is to be secured.

The present fuel system also advantageously provides a modular, one-piece fuel pump housing including a plurality of outwardly opening pump cavities, a radially enclosed drive shaft, a pump head engagement surface, and a plurality of follower guide surfaces in respective pump cavities, wherein the follower guide surfaces, the head engagement surface, and the drive shaft len mounting surfaces are the only surfaces that require precise machining to provide proper alignment between the drive shaft and the cooperating fuel pumping elements of the pump. By providing a pump head secured to the pump housing opposite the drive shaft and a pump unit retained in the pump head by a bracket that causes the pump unit to extend into the pump cavity of the pump housing in a spaced, non-contacting relationship with the pump housing, the present invention further enables the pump unit to be relatively easily removed and replaced to provide inexpensive maintenance of the pump assembly and/or the ability to change pump units to adjust the effective displacement of the fuel pump assembly.

Furthermore, the fuel system of the present invention minimizes the number of fuel seepage points by reducing system components and providing fail-safe, redundant low pressure fuel outlets throughout the system to capture and return to the fuel system any fuel that may leak through primary sealing areas. Likewise, the present fuel system may include both two pump control valves and two injection control valves so that one corresponding valve can take over should the other corresponding valve fail.

The present invention also provides an improved accumulator having a labyrinth of interconnected chambers, the chambers being elongated and cylindrical and arranged in a generally parallel relationship and intersecting a vertical plane through the accumulator housing in a two-dimensional array. The accumulator chambers are particularly oriented to minimize the physical dimensions of the accumulator housing while being dimensioned to provide a minimum total volume sufficient to prevent a fuel pressure drop of more than five percent during each injection event, depending on factors such as the compressibility of the fuel, the operating pressure of the fuel, the maximum possible required injection volumes, the timing range and the injection duration required for the engine, the maximum effective displacement of each pump unit, the fuel leakage from the system, the compression of the fuel in the fuel lines, and the fuel loss to the outlet during valve element movement between the fully open and fully closed positions.

The disclosed invention provides a number of additional features such as (1) the integration of a rotary pump and distributor with a single drive shaft arrangement; (2) the provision of a distributor with axially slidable plunger regulators in combination with a dedicated injection control valve; (3) the provision of a range of pump head/accumulator designs to accommodate pump control valves and check valves; (4) the provision of an ultra-compact pump head and integrated pump chamber designs; (5) the provision of a transversely oriented pump control valve to reduce the volume trapped in the accumulator to an absolute minimum; (6) the provision of a pump unit and a transversely oriented pump control valve mounted in the cylinder of the pump unit; (7) various accumulator designs to simplify the formation and manufacture of the accumulator; (8) the provision of a separately mounted accumulator; (9) the provision of various edge filter mounting concepts for use in the disclosed fuel system; and (10) the provision of quantity regulation and cavitation control devices in the disclosed fuel system.

Industrial applicability

The compact, heavy-duty fuel system of the present invention and its components can be used in a variety of internal combustion engines of a vehicle or industrial machine requiring accurate and reliable high pressure fuel delivery. However, the heavy-duty fuel system of the present invention is particularly useful in small and medium displacement diesel truck engines and can be particularly adapted to existing diesel engine designs without significant engine modifications.

Claims (72)

1. Electronically controllable high-pressure fuel pump assembly (10, 400, 486, 988) for supplying fuel at a predetermined pressure through a plurality of fuel injection lines (806, 1324, 1452, 1612, 1660) to corresponding cylinders of a multi-cylinder internal combustion engine, the pump assembly (10, 400, 486, 988) comprising: a) a modular or unitary assembly (10, 400, 486, 988) adapted to be attached to the engine, the modular or unitary assembly (10, 400, 486, 988) comprising: i. Pump means for pressurizing fuel above the predetermined pressure, the pump means comprising a pump housing (22, 428, 500) having a fastening means for fastening the modular or unitary assembly (10, 400, 486, 988) to the engine, a pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) and a pump plunger (122, 416, 418, 490, 492, 752, 884, 886, 1004) inserted into the pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) for reciprocating movement between an advanced position and a retracted position, ii. a storage means for collecting and temporarily storing fuel received from the pump means at high pressure, the storage means comprising a storage housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) containing at least one storage chamber (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232), the storage housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) being attached to the pump housing (22, 428, 500), and iii. a fuel distribution means for enabling sequential periodic fluid communication between the storage chamber (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) and the engine cylinders, wherein the distributor means comprises a distributor housing (44, 410, 528, 713, 782, 1264, 1402) attached to the pump housing (22, 428, 500); b) a first solenoid-operated pump control valve (18, 766, 994, 997, 1076, 1100, 1102) for controlling the pump means to maintain a desired fuel pressure in the storage chamber (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232), the first solenoid-operated pump control valve (18, 19, 766, 994, 997, 1076, 1100, 1102) being attached to the modular assembly (10, 400, 486, 988) and for controlling an effective displacement of the pump plunger (122, 416, 418, 490, 492, 752, 884, 886, 1004), wherein the effective displacement is defined in part by the advanced position of the pump plunger (122, 416, 418, 490, 492, 752, 884, 886, 1004); and c) a first solenoid-operated injection control valve (20, 440, 440, 454, 1406, 1604, 1634) for controlling the timing and amount of fuel injected into each engine cylinder depending on engine operating conditions, the first solenoid-operated injection control valve (20, 440, 440, 454, 1406, 1604, 1634) being attached to the modular or unitary assembly (10, 400, 486, 988). 2. Pump assembly according to claim 1, characterized in that the pump assembly (10, 400, 486, 988) further comprises a second electromagnetically actuated injection control valve (20, 440, 440, 454, 1406, 1604, 1634) for controlling the timing and amount of injection into each engine cylinder even if the first electromagnetically actuated injection control valve (20, 440, 440, 454, 1406, 1604, 1634) fails. 3. Pump arrangement according to claim 1 or 2, characterized in that the pump means comprises a plurality of pump chambers (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) and a plurality of pump plungers (122, 416, 418, 490, 492, 752, 884, 886, 1004) which are arranged in the pump pen chambers (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) are installed for reciprocating movement, and wherein the arrangement (10, 400, 486, 988) further comprises a plurality of electromagnetically operated pump control valves (18, 19, 766, 994, 997, 1076, 1100, 1102), the number of which corresponds to that of the pump chambers, wherein the electromagnetically operated pump control valves (18, 19, 766, 994, 997, 1076, 1100, 1102) are each connected to the pump chambers (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) for controlling the effective displacement of each associated pump plunger. 4. Pump arrangement according to one of claims 1 to 3, characterized in that the pump arrangement (10, 400, 486, 988) further comprises means for generating a pressure signal which indicates the pressure of the fuel in the storage means, and control means for controlling the electromagnetically actuated pump control valves (18, 19, 766, 994, 997, 1076, 1100, 1102) for adjusting the effective displacement of the pump plungers (122, 416, 418, 490, 492, 752, 884, 886, 1004) in dependence on the pressure signal so that the pressure of the fuel in the storage means becomes equal to the predetermined pressure. 5. Pump assembly according to one of claims 1 to 4, characterized in that the fuel distribution means comprises a plurality of injection line valves (784) for controlling the flow of fuel to corresponding cylinders through corresponding fuel injection lines (806), each of the injection line valves (784) comprising a slide valve element (808) which is reciprocally mounted in the distributor housing (782). 6. Pump assembly according to claim 5, characterized in that the fuel distribution means further comprises a distributor camshaft (786) rotatably mounted in the distributor housing (782), the distributor camshaft (786) having at least one cam (796) which causes the distributor slide valve elements (808) to reciprocate when the distributor camshaft (786) is rotated, the slide valve elements (808) being adapted to reciprocate along corresponding axial lines are installed, which are parallel to the axis of rotation of the distributor camshaft (786). 7. Pump assembly according to claim 6, characterized in that each of the plurality of slide valve elements (808) is movable to an open position to define a corresponding fuel injection period in which high pressure fuel can flow to the corresponding engine cylinder via the corresponding fuel injection line (806) and to a closed position in which the flow of fuel through the corresponding fuel injection line (806) is blocked, each of the plurality of injection line valves (784) being spool-like and having a protruding or thickened portion (812) formed on the slide valve element (808) for blocking the flow of fuel when the corresponding injection line valve (784) is in the closed position. 8. A pump assembly according to claim 7, characterized in that the sliding valve element (808) has a cylindrical portion with a first end and a second end, an annular groove (814) being formed in the cylindrical portion adjacent the protruding or thickened part (812) so that fuel can flow to the engine cylinders when the corresponding injection line valve (784) is in the open position, the pump assembly further comprising a clamping means arranged adjacent to the first end for clamping the second end into engagement with the at least one cam (796). 9. Pump arrangement according to one of the preceding claims, characterized in that the pump arrangement (10, 400, 486, 988) is compact, that the pump housing (22, 428, 500) has minimal dimensions in mutually perpendicular lateral, radial and axial directions, wherein the pump housing (22, 428, 500) has at least one pump cavity (28, 30, 876, 878, 996) with a first pump axis extending in the radial direction and a drive shaft cavity (24, 434, 558) adjacent to or adjacent to one end of the pump pen cavity (28, 30, 876, 878, 996) having a drive axis extending in the axial direction, wherein a drive shaft (26, 756, 888) is arranged in the drive shaft cavity (24, 434, 558) for rotation about the drive axis, wherein the pump plunger (122, 416, 418, 490, 492, 752, 884, 886, 1004) is installed in the pump cavity (28, 30, 876, 878, 996) for reciprocating movement along the first pump axis depending on the rotational movement of the drive shaft (26, 756, 888), and wherein the accumulator housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) includes at least one elongated storage chamber (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) for collecting and temporarily storing fuel at high pressure and is attached to the pump housing (22, 428, 500) adjacent to the other end of the pump cavity (28, 30, 876, 878, 996), the central axis of the elongated storage chamber (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) being arranged parallel to the drive axis. 10. Pump arrangement according to claim 9, characterized in that the accumulator housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) has an axial dimension which is substantially greater than the axial dimension of the pump housing (22, 428, 500), whereby an axial overhang (38, 987) of the accumulator housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) is created with respect to the pump housing (22, 428, 500). 11. Pump assembly according to claim 9 or 10, characterized in that the pump housing (22, 428, 500) contains at least one additional pump cavity (28, 30, 876, 878, 996) with a second pump axis that runs parallel to the first pump axis and perpendicular to the drive shaft, and that the pump assembly (10, 400, 486, 988) further comprises a second pump plunger (122, 416, 418, 490, 492, 752, 884, 886, 1004) that is installed for reciprocating movement along the second pump axis dependent on the rotational movement of the drive shaft (26, 756, 888). 12. Pump arrangement according to claim 10 or 11, characterized in that the distributor housing (44, 410, 528, 713, 782, 1264, 1402) is the pump housing (22, 428, 500) adjacent to the drive shaft cavity (24, 434, 558) in a spaced, substantially parallel relationship to the axial overhang (38, 987) of the accumulator housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244). 13. Pump assembly according to claim 12, characterized in that the distributor housing (44, 410, 528, 713, 782, 1264, 1402) contains a rotor bore (214, 475) and the distributor means further comprises a distributor rotor (216, 425, 462, 708, 1424) which is installed for rotation in the rotor bore (214, 475), the rotor (216, 425, 462, 708, 1424) being driven for rotation by the drive shaft (26, 756, 888), the rotor (216, 425, 462, 708, 1424) containing an axial supply passage (226, 438), which is in fluid communication for receiving fuel from the storage chamber (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232), the rotor (216, 425, 462, 708, 1424) also including a first radial supply passage (230) which is in fluid communication with the axial supply passage (226, 438), the distributor housing (44, 410, 528, 713, 782, 1264, 1402) including a group of receiving openings (231) which are adapted for connection to the respective engine cylinders through respective fuel injection lines (806, 1324, 1452, 1612, 1660), the Receiving openings (231) are spaced around the circumference of the rotor (216, 425, 462, 708, 1424), the group of receiving openings (231) being arranged in positions which are sequentially aligned with the first radial supply passage (230) as the rotor (216, 425, 462, 708, 1424) is rotated to define individual separate periods during each rotation of the rotor (216, 425, 462, 708, 1424) in which the corresponding engine cylinders can be brought into fluid communication with the storage chamber (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232), and preferably wherein the axis of rotation of the rotor (216, 425, 462, 708, 1424) is coaxial with the axis of rotation of the drive shaft (26, 756, 888) or wherein the axis of rotation of the rotor (216, 425, 462, 708, 1424) is perpendicular to the axis of rotation of the drive shaft (26, 756, 888). 14. Pump assembly according to claim 13, characterized in that the pump assembly (10, 400, 486, 988) further comprises a fuel supply line for fluidly connecting the axial supply passage (226, 438) to the storage chamber (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232), the supply line having a supply opening for conducting fuel from the storage chamber (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) to the rotor bore (214, 475), the supply opening being arranged in a supply plane which is perpendicular to the axis of rotation of the rotor and axially from the group of receiving openings (231), wherein the rotor (216, 425, 462, 708, 1424) includes a radial receiving passage (242, 712) arranged axially in the supply plane, and preferably wherein the distributor housing (44, 410, 528, 713, 782, 1264, 1402) includes a distributor housing outlet opening arranged at one end of the rotor bore (214, 475) for connection to a low pressure outlet (284), wherein the rotor (216, 425, 462, 708, 1424) includes a first axial outlet passage (286) which is in fluid communication with the distributor housing outlet opening. 15. A pump assembly according to claim 14, characterized in that the rotor (216, 425, 462, 708, 1424) further includes a first radial discharge passage communicating with an axial discharge passage (286) and with a first discharge groove formed in either the rotor (216, 425, 462, 708, 1424) or the rotor bore (214, 475) axially disposed between the first radial supply passage (230) and the radial receiving passage (242, 712) for collecting any fuel leaking through the tight fit gap between the rotor (216, 425, 462, 708, 1424) and the rotor cavity located between the radial supply passage (230) and the radial receiving passage (242, 712). 16. A pump assembly according to any one of claims 12 to 15, characterized in that it further comprises a pressure booster pump means disposed between the distributor means and the pump housing (22, 428, 500) for receiving fuel from a force fuel source and for supplying fuel to the pump cavity (28, 30, 876, 878, 996) at a pressure sufficient to deliver an adequate amount of fuel to the pump cavity (28, 30, 876, 878, 996) throughout the operating range of the engine, and preferably wherein the pressure boost pump means comprises a shaft extension coupled to the drive shaft (26, 756, 888) of the fuel pump at one end and to the distributor rotor (216, 425, 462, 708, 1424) at the other end, the distributor housing (44, 410, 528, 713, 782, 1264, 1402) having a sealing recess which seals the end of the distributor rotor (216, 425, 462, 708, 1424) adjacent to or adjacent to the shaft extension. 17. Pump assembly according to claim 13, characterized in that the rotor (216, 425, 462, 708, 1424) includes a pressure equalization groove extending a sufficient circumferential distance around the rotor (216, 425, 462, 708, 1424) at an axial position to fluidly connect all of the receiving openings (231) except the receiving opening (231) which is in fluid communication with the first radial supply passage (230), and preferably wherein the receiving openings (231) are equiangularly spaced around the circumference of the rotor (216, 425, 462, 708, 1424) to maximize the space between the receiving openings (231), and again preferably wherein the distribution means comprises a Supply groove (244) contained either in the rotor (216, 425, 462, 708, 1424) or the rotor bore (214, 475), the supply groove (244) being arranged so that it is in communication at all times with the radial receiving passage (242, 712) of the rotor and the fuel supply line. 18. Pump arrangement according to one of claims 14 to 17, characterized in that the distributor means comprises an injection control means for controlling the timing and amount of fuel injected into each engine cylinder depending on engine operating conditions, the injection control means comprising a first solenoid injection control valve (20, 440, 440, 454, 1406, 1604, 1634) which is attached to the distributor housing (44, 410, 528, 713, 782, 1264, 1402) and arranged to control the flow of fuel through the fuel supply line, wherein the first solenoid injection control valve (20, 440, 440, 454, 1406, 1604, 1634) is a three-way valve which, when energized, is operable to connect the axial supply passage (226, 438) of the rotor (216, 425, 462, 708, 1424) to the storage means, and when de-energized, is operable to connect the axial supply passage of the rotor bore (10, 400, 486, 988)4 to a low pressure outlet, wherein the distributor housing (44, 410, 528, 713, 782, 1264, 1402) has an elongated first valve cavity (270, 460) for Receptacle for the first solenoid injection control valve (20, 440, 440, 454, 1406, 1604, 1634). 19. Pump arrangement according to claim 18, characterized in that the injection control means comprises a second solenoid injection control valve (20, 440, 440, 454, 1406, 1604, 1634) which is attached to the distributor housing (44, 410, 528, 713, 782, 1264, 1402) and arranged parallel to the first solenoid injection control valve (20, 440, 440, 454, 1406, 1604, 1634) for controlling the flow of fuel through the fuel supply line, the second solenoid injection control valve (20, 440, 440, 454, 1406, 1604, 1634) being a three-way valve which, in the energized state, is arranged to connect the axial supply passage (226, 438) of the rotor (216, 425, 462, 708, 1424) with the storage means, and in the off state is operable to connect the axial supply passage of the rotor (216, 425, 462, 708, 1424) with a low pressure outlet, wherein the distributor housing (44, 410, 528, 713, 782, 1264, 1402) contains a second valve cavity (272, 460) with a central axis which is parallel to a central axis of the first valve cavity (270, 460), the central axes lying in the supply plane which contains the radial supply passage (230) which supplies fuel to the axial supply passage (226, 438) of the rotor (216, 425, 462, 708, 1424), wherein the first and second cavities (270, 272, 460) are arranged on opposite sides of the rotor (216, 425, 462, 708, 1424), and wherein preferably the first and second valve cavities (270, 272, 460) are connected by a rotor supply bore (276, 448) having a central axis lying in the supply plane, wherein the Rotor cavity supply port is in fluid communication with the rotor supply bore (276, 448), the manifold means including a two-way check valve (453) disposed in the rotor supply bore (276, 448) to prevent fuel supplied from one valve cavity (270, 272, 460) from flowing into the other valve cavity (270, 272, 460). 20. Pump arrangement according to one of the preceding claims, characterized in that the pump arrangement (10, 400, 486, 988) is an ultra high pressure fuel pump arrangement, that the pump arrangement (10, 400, 486, 988) is modular or designed as a unit and is designed for attachment to the internal combustion engine, that the multi-cylinder internal combustion engine has a predetermined operating range and has reciprocating pistons which are assigned to the corresponding cylinders, that the storage means is a high pressure accumulator (12, 402, 466, 702, 776, 1168, 1208, 1606, 1562) for storing a predetermined volume of fuel at the predetermined pressure, and that the fuel distribution means for establishing a sequential periodic fluid connection between the storage means and the engine cylinders through the fuel injection lines (806, 1324, 1452, 1612, 1660) associated with the respective engine cylinders to effect a periodic injection of fuel into the respective engine cylinder in synchronism with the movement of the piston in the respective engine cylinder, the high pressure storage means comprising a high strength, compact storage housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) containing a fluidically connected labyrinth of storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) having a sufficient total volume so that controlled amounts of fuel at the operating pressure can be delivered to each engine cylinders at appropriate times throughout the operating range of the engine as determined by the fuel distribution means. 21. Pump arrangement according to claim 20, characterized in that the pumping means comprises at least one pump unit (86, 88, 880, 882, 992, 993, 1112, 1114) responsive to a control signal to vary the amount of fuel pumped, and further comprising pressure sensing means for determining the pressure in the storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232), and pump control means for generating the pump control signal to maintain the fuel pressure in the storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) at the predetermined operating pressure, and preferably wherein the storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) are elongated and cylindrical and are connected by connecting passages. 22. Pump arrangement according to claim 21, characterized in that the storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) are arranged next to one another and are aligned in a substantially parallel relation to one another. 23. Pump arrangement according to claim 21, characterized in that the storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) are arranged so that they intersect a vertical plane through the storage housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) in a two-dimensional arrangement. 24. Pump arrangement according to claim 23, characterized in that the two-dimensional arrangement has an upper row (54) of four storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) and a lower row (56) of three storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232). 25. Pump arrangement according to claim 21, characterized in that the storage housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) is formed from a whole one-piece block, and wherein the storage means comprises a plurality of plugs arranged at the ends of the respective storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) for the fluid-tight closure of the ends of the storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232), and preferably wherein the pump housing (22, 428, 500) includes a plurality of pump cavities (28, 30, 876, 878, 996) and the reservoir housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) includes a plurality of pump unit recesses (82, 84, 1116, 1118) aligned and communicating with the corresponding pump cavities (28, 30, 876, 878, 996), and wherein the pump means includes a plurality of pump units (86, 88, 880, 882, 992, 993, 1112, 1114), each pump unit (86, 88, 880, 882, 992, 993, 1112, 1114) is arranged in a corresponding pump cavity (28, 30, 876, 878, 996) and the associated pump unit recess (82, 84, 1116, 1118), and further preferably wherein each pump unit (86, 88, 880, 882, 992, 993, 1112, 1114) has a pump cylinder (116, 820, 844, 890, 892, 995) which has a pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) and a pump plunger (122, 416, 418, 490, 492, 752, 884, 886, 1004) mounted for reciprocating movement in the pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002), and preferably further comprising a pump unit control valve (18, 19, 766, 994, 997, 1076, 1100, 1102) mounted to the pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) for controlling the effective displacement of the pump plunger (122, 416, 418, 490, 492, 752, 884, 886, 1004), wherein the reservoir housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) has a first side for engagement with the pump housing (22, 428, 500) and a second side formed opposite the first side, wherein the pump unit control valve (18, 19, 766, 994, 997, 1076, 1100, 1102) is arranged on the second side of the reservoir housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) directly opposite a corresponding pump unit recess (82, 84, 1116, 1118). 26. Pump arrangement according to claim 25, characterized in that the pump means comprises a camshaft (26, 756, 888) which is rotatably installed in the pump housing (22, 428, 500), the camshaft (26, 756, 888) having a plurality of cams (172) which each set the plungers (122, 416, 418, 490, 492, 752, 884, 886, 1004) in a reciprocating movement when the camshaft (26, 756, 888) is rotated. 27. Pump arrangement according to claim 26, characterized in that the pump means comprises a plurality of tappet or driver arrangements (148) which are respectively associated with the pump units (86, 88, 880, 882, 992, 993, 1112, 1114), each driver arrangement (148) being installed for reciprocating movement in the pump cavity (28, 30, 876, 878, 996) in which the corresponding pump unit (86, 88, 880, 882, 992, 993, 1112, 1114) is installed, and with the plunger (122, 416, 418, 490, 492, 752, 884, 886, 1004) of the corresponding pump unit (86, 88, 880, 882, 992, 993, 1112, 1114), and wherein the pump means comprises a driver tension spring (158, 898, 900) for tensioning the driver assembly (148) into engagement with a corresponding cam (172) on the camshaft, so that the driver assembly (148) and the connected plunger (122, 416, 418, 490, 492, 752, 884, 886, 1004) are set in a reciprocal movement when the camshaft (26, 756, 888) is rotated, and preferably wherein each cam (172) has at least one nose or elevation (204) which causes an associated plunger (122, 416, 418, 490, 492, 752, 884, 886, 1004) to undergo a forward stroke and a reverse stroke for each revolution of the camshaft (26, 756, 888), the total number of lobes (204) on all cams (172) being selected to effect a forward stroke for each of the periodic injections into each of the engine cylinders. 28. Pump arrangement according to one of claims 25 to 27, characterized in that each pump unit (86, 88, 880, 882, 992, 993, 1112, 1114) has a pump holder (104, 998) which surrounds the cylinder (116, 820, 844, 890, 892, 995) in order to support the pump unit (86, 88, 880, 882, 992, 993, 1112, 1114) in the corresponding pump unit recess (82, 84, 1116, 1118) of the storage housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244), each pump unit (86, 88, 880, 882, 992, 993, 1112, 1114) extending into the corresponding pump cavity (28, 30, 876, 878, 996) without being in direct contact with the pump housing (22, 428, 500). 29. Pump arrangement according to claim 28, characterized in that each pump unit (86, 88, 880, 882, 992, 993, 1112, 1114) contains a pump unit inlet which is in communication with a fuel source for supplying fuel into the pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002), and a pump unit outlet which is in communication with the labyrinth of storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232), each pump unit (86, 88, 880, 882, 992, 993, 1112, 1114) has a pump unit check valve (136, 824, 854, 970, 1014, 1128) which only allows fuel flow in one direction from the pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) through the pump unit outlet into the storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232). 30. Pump assembly according to claim 29, characterized in that each pump unit check valve (136, 824, 854, 970, 1014, 1128) has a check valve recess (132) contained in the accumulator housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) for forming a fluid communication path between a corresponding disk outlet passage (130) and the accumulator chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232), each pump unit check valve (136, 824, 854, 970, 1014, 1128) further comprises a check valve element which is designed to be tensioned into a closed position by the pressure of the fuel in the storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) until the pressure of the fuel in the corresponding pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) exceeds the pressure in the storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232), at which time the check valve element is opened so that fuel can flow from the corresponding pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) and through the check valve recess (132) into the storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232). 31. Pump arrangement according to claim 29 or 30, characterized in that each pump unit (86, 88, 880, 882, 992, 993, 1112, 1114) has a Disc (112) arranged in the holder (104, 998) at one end of the cylinder (116, 820, 844, 890, 892, 995) for closing the corresponding pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002), wherein the pump unit disc (112) contains the pump unit inlet and the pump unit outlet, and wherein the holder (104, 998) is inserted into the corresponding pump unit recess (82, 84, 1116, 1118) of the storage housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) to clamp the cylinder (116, 820, 844, 890, 892, 995) and the disc (112) in an axially stacked relationship against the accumulator housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244), the pump unit outlet having a disc outlet passage (130) located in the center of the disc (112), the pump unit inlet having an annular disc groove (126) concentrically located on one side of the disc (112), and at least one axial disc inlet passage (128) from the pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) to the annular disc groove (126). 32. Pump arrangement according to one of claims 29 to 31, characterized in that the accumulator housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) contains at least one common fuel supply passage (90) for supplying fuel to all pump units (86, 88, 880, 882, 992, 993, 1112, 1114) and a plurality of fuel supply side channels (78, 80) each extending between the common fuel supply passage (90) and the pump unit recesses (82, 84, 1116, 1118), each fuel supply side channel (78, 80) being connected at one end to the annular Disc groove (126) formed in the corresponding pump unit recess (82, 84, 1116, 1118) and at the other end communicates with the common fuel supply passage (90). 33. Pump arrangement according to claim 32, characterized in that the pump arrangement (10, 400, 486, 988) further comprises a plurality of pump unit control valves (18, 19, 766, 994, 997, 1076, 1100, 1102) which are respectively associated with the fuel supply side channels (78, 80) for controlling the flow of fuel through the corresponding fuel supply side channels (78, 80) in response to a pump unit control signal in order to control the amount of fuel pumped by the corresponding pump unit (86, 88, 880, 882, 992, 993, 1112, 1114) into the storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) during each reciprocating cycle of the corresponding pump plunger (122, 416, 418, 490, 492, 752, 884, 886, 1004), and which preferably further comprises a pressure sensing means for determining the pressure in the storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) and a pump unit valve control means for generating the pump unit control signal for each pump unit control valve (18, 19, 766, 994, 997, 1076, 1100, 1102) to maintain the fuel pressure in the storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) at the predetermined operating pressure. 34. Pump assembly according to one of claims 20 to 33, characterized in that the accumulator housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) contains an accumulator outlet passage which communicates with each pump unit recess (82, 84, 1116, 1118) and with the common fuel supply passage (90), each pump unit (86, 88, 880, 882, 992, 993, 1112, 1114) having a pump unit outlet means for directing fuel discharged from the pump unit (86, 88, 880, 882, 992, 993, 1112, 1114) into the accumulator outlet passage, each pump unit outlet means further comprising a recess space (138) formed between the respective bracket (104, 998) and the respective pump unit recess (82, 84, 1116, 1118), each recess space (138) communicating with the respective accumulator outlet passage. 35. Pump assembly according to claim 34, characterized in that the outlet means further comprises a pump unit gap (140) between the respective cylinder (116, 820, 844, 890, 892, 995) and the bracket (104, 998), an outlet groove (144) formed on the Surface of the corresponding pump plunger (122, 416, 418, 490, 492, 752, 884, 886, 1004) and a support outlet passage (142) communicating with the pump unit space (140) at all times and intermittently communicating with the outlet groove (144) during the reciprocating movement of the corresponding pump plunger (122, 416, 418, 490, 492, 752, 884, 886, 1004), whereby fuel discharged from the corresponding pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) can be directed between the corresponding cylinder (116, 820, 844, 890, 892, 995) and the pump plunger (122, 416, 418, 490, 492, 752, 884, 886, 1004) is collected in the discharge groove (144) for intermittent discharge through the corresponding discharge passage, and preferably wherein each pump unit gap (140) is in fluid communication to receive fuel leaking from the contact surface between the corresponding disc (112) and the bracket (104, 998), and wherein each recess gap (138) is adapted to receive fuel leaking from the contact surface between the corresponding disc (112) and the accumulator housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) so that fuel leaking from the contact surfaces can be returned to the common fuel supply passage (90), and preferably wherein the fuel flows from the corresponding pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) through the corresponding disc outlet passage (130) and the check valve recess (132) into the storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232). 36. Pump arrangement according to claim 20, characterized in that the pump means comprises a pump housing (22, 428, 500) and a pump head (1170, 1189) which is fastened to the pump housing, the storage housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) being arranged at a distance from the pump head (1170, 1189), and further comprising a first delivery pipe (1194) for fluidly connecting the pump means to the storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232). 37. A pump assembly according to any preceding claim, characterized in that the pump means is designed to deliver fuel above about 34.47 MPa (5000 psi), the pump means comprising a pump housing (22, 428, 500) containing at least one pump cavity (28, 30, 876, 878, 996) opening into a head engagement surface (32), the storage means being a high pressure storage means in fluid communication with the pump means for accumulating a predetermined volume of fuel at a predetermined operating pressure above about 34.47 MPa (5000 psi), the high pressure storage means comprising a high strength compact storage housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) containing at least one storage chamber (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) and secured in contact with the head engagement surface (32) of the pump housing (22, 428, 500) to form an end wall for the pump cavity (28, 30, 876, 878, 996). 38. Pump arrangement according to claim 37, characterized in that the storage housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) contains a fluidically interconnected labyrinth of storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232), the total volume of which is sufficient so that a controlled amount of fuel at the predetermined operating pressure can be delivered to each engine cylinder at suitable times during the entire operating range of the engine, the storage housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) is formed from a one-piece block containing the labyrinth of storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) which is constructed and arranged to form surrounding walls which are sufficiently strong to withstand the forces generated when the storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) are filled with fuel at the predetermined operating pressure. 39. Pump arrangement according to one of the preceding claims, characterized in that the pump housing (22, 428, 500) is compact and has minimal dimensions in mutually perpendicular lateral, radial and axial directions, wherein the pump housing (22, 428, 500) includes at least one pump cavity (28, 30, 876, 878, 996) having a first central axis extending in the radial direction, wherein the pump plunger (122, 416, 418, 490, 492, 752, 884, 886, 1004) is installed in the pump cavity (28, 30, 876, 878, 996) for reciprocating movement along the first central axis, wherein the accumulator housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) is mounted on the pump housing (22, 428, 500) is mounted adjacent to one end of the pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002), wherein at least one of the axial dimension and the lateral dimension of the accumulator housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) is greater than the corresponding dimension of the pump housing (22, 428, 500), whereby an axial overhang (38, 40, 987) of the accumulator housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) with respect to the pump housing (22, 428, 500), and wherein the fuel distribution means comprises a distributor body cantileveredly mounted to the pump housing (22, 428, 500) in a parallel, substantially spaced relationship with respect to the overhang (38, 40, 987) of the accumulator housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244). 40. Pump arrangement according to claim 39, characterized in that the first solenoid injection control valve (20, 440, 440, 454, 1406, 1604, 1634) is attached to the distributor housing (44, 410, 528, 713, 782, 1264, 1402) in the space between the distributor housing (44, 410, 528, 713, 782, 1264, 1402) and the projecting overhang (38, 40, 987) of the accumulator housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244), and preferably wherein the injection control means comprises a second solenoid injection control valve (20, 440, 440, 454, 1406, 1604, 1634) for controlling the flow of fuel from the storage means to the respective engine cylinders, wherein the second solenoid injection control valve (20, 440, 440, 454, 1406, 1604, 1634) is mounted on the distributor housing (44, 410, 528, 713, 782, 1264, 1402) adjacent to the first solenoid injection control valve (20, 440, 440, 454, 1406, 1604, 1634) in the space between the distributor housing housing (44, 410, 528, 713, 782, 1264, 1402) and the cantilevered overhang (38, 40, 987) of the storage housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244). 41. Pump arrangement according to one of the preceding claims, characterized in that the pump housing (22, 428, 500) contains an outwardly opening pump cavity (28, 30, 876, 878, 996), a drive shaft (26, 756, 888) is rotatably installed in the pump housing (22, 428, 500), a pump head (1170, 1189) can be fastened to the pump housing (22, 428, 500) in order to close the outwardly opening pump cavity (28, 30, 876, 878, 996), the pump head (1170, 1189) containing a pump unit recess (82, 84, 1116, 1118) which in connection with the pump cavity (28, 30, 876, 878, 996), wherein a replaceable pump unit (86, 88, 880, 882, 992, 993, 1112, 1114) is provided, which has a pump cylinder (116, 820, 844, 890, 892, 995) which contains the pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) and the pump plunger (122, 416, 418, 490, 492, 752, 884, 886, 1004) which is adapted to be mounted for reciprocating movement in the pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002), depending on the rotation of the drive shaft (26, 756, 888), wherein the replaceable pump unit (86, 88, 880, 882, 992, 993, 1112, 1114) has a mounting means for fastening the pump unit (86, 88, 880, 882, 992, 993, 1112, 1114) in the pump unit recess (82, 84, 1116, 1118) of the pump head (1170, 1189) in a position so that it is at least partially inserted into the pump cavity (28, 30, 876, 878, 996) in a spaced, non-contacting relationship with the pump housing (22, 428, 500), and preferably wherein the pump housing (22, 428, 500) has a plurality of the outwardly opening pump cavities (28, 30, 876, 878, 996), the pump head (1170, 1189) includes a plurality of the pump unit recesses (82, 84, 1116, 1118) each arranged for communication with the pump cavities (28, 30, 876, 878, 996), and further comprising a plurality of the interchangeable pump units (86, 88, 880, 882, 992, 993, 1112, 1114), each pump unit (86, 88, 880, 882, 992, 993, 1112, 1114) a pump cylinder (116, 820, 844, 890, 892, 995) defining a pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002), a pump plunger (122, 416, 418, 490, 492, 752, 884, 886, 1004) incorporated in the pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) to reciprocate when the drive shaft (26, 756, 888), and a mounting means for securing the pump unit (86, 88, 880, 882, 992, 993, 1112, 111-4) in a corresponding pump unit recess (82, 84, 1116, 1118) of the pump head (1170, 1189) in a position such that it extends at least partially into the pump cavity (28, 30, 876, 878, 996) in a spaced, non-contacting relationship with the pump housing (22, 428, 500). 42. Pump arrangement according to claim 41, characterized in that the pump head (1170, 1189) forms at least a partial end wall (114) for the pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002), wherein the pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) is arranged immediately next to the pump head (1170, 1189), and/or wherein the pump cylinder (116, 820, 844, 890, 892, 995) is a one-piece construction which has an inner end which rests or System with the pump head (1170, 1189). 43. Pump arrangement according to claim 42, characterized in that the pump cylinder (116, 820, 844, 890, 892, 995) has a pump inlet passage which is designed to be connected to a fuel source in order to conduct fuel into the pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) and a pump outlet passage through which fuel can be discharged from the pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002), and wherein the pump unit (86, 88, 880, 882, 992, 993, 1112, 1114) has a pump unit check valve (136, 824, 854, 970, 1014, 1128) which is arranged at least partially in the pump outlet passage so that only a fuel flow in one direction from the pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) through the pump outlet passage is possible, wherein the pump safety check valve (136, 824, 854, 970, 1014, 1128) has a check valve seat which is formed on the pump cylinder (116, 820, 844, 890, 892, 995). 44. A pump assembly according to claim 41, characterized in that the pump head (1170, 1189) has a discharge passage for receiving high pressure fuel from the pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002), the pump cylinder (116, 820, 844, 890, 892, 995) having an inner end arranged in abutment with the pump head (1170, 1189) to form a high pressure connection exposed to the high pressure fuel discharged from the pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) is delivered to the delivery passage, the high pressure connection being the only connection disposed between the pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) and the delivery passage that is exposed to the high pressure fuel. 45. Pump arrangement according to claim 41, characterized in that the pump arrangement (10, 400, 486, 988) further comprises a plurality of pump unit control valves (18, 19, 766, 994, 997, 1076, 1100, 1102) respectively associated with the pump chambers (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) for controlling the amount of high pressure fuel that is discharged from the corresponding pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) through a corresponding pump plunger (122, 416, 418, 490, 492, 752, 884, 886, 1004) and having a valve cavity formed in each of the pump cylinders (116, 820, 844, 890, 892, 995), wherein each of the plurality of pump unit control valves (18, 19, 766, 994, 997, 1076, 1100, 1102) has a control valve element installed for reciprocating movement in a corresponding valve cavity, and preferably, wherein each of the plurality of pump unit control valves (18, 19, 766, 994, 997, 1076, 1100, 1102) has an annular valve seat formed on the corresponding pump cylinder (116, 820, 844, 890, 892, 995) in the valve cavity. 46. Pump arrangement according to one of the preceding claims, characterized in that the storage means comprises a high-strength, compact storage housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) containing a fluidically connected labyrinth of storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232), the total volume of which is sufficient so that a controlled amount of fuel at a predetermined operating pressure can be delivered to each engine cylinder at appropriate times throughout the entire operating range of the engine, the storage housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) is formed from a one-piece block containing the labyrinth of storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) designed and arranged to form surrounding walls sufficiently strong to withstand the forces generated when the storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) are filled with fuel at the predetermined operating pressure, the storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) being arranged are such that they intersect a vertical plane through the storage housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) in at least one two-dimensional arrangement, and preferably wherein the storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) are formed by drilling the one-piece block, and wherein the storage means comprises a plurality of plugs for closing the storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232). 47. Pump arrangement according to claim 46, characterized in that the two-dimensional arrangement has an upper row (54) of a plurality of storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) and a lower row (56) of a plurality of storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232), and preferably wherein the storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) are fluidically connected to one another by a first transverse passage which connects the upper row (54) of storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) and a second transverse Passage which intersects the lower row (56) of storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232). 48. Pump arrangement according to claim 47, characterized in that the storage housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) has a plurality of pump recesses (82, 84, 1116, 1118), the storage means further comprising a plurality of pump units (86, 88, 880, 882, 992, 993, 1112, 1114) which are each received in the pump recesses (82, 84, 1116, 1118) and are carried or supported by the storage housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244). are held, wherein each of the pump unit recesses (82, 84, 1116, 1118) is in fluid communication with the storage chambers (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232), and preferably wherein the storage housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) contains at least one common fuel supply passage (90) for supplying fuel to all pump units (86, 88, 880, 882, 992, 993, 1112, 1114) and a plurality of fuel supply side channels (78, 80) each extending between the common fuel supply passage (90) and the pump unit recesses (82, 84, 1116, 1118), each fuel supply side channel (78, 80) communicating at one end with the corresponding pump unit recess (82, 84, 1116, 1118) and at the other end communicating with the common fuel supply passage (90). 49. Pump arrangement according to one of the preceding claims, characterized in that the pump arrangement (10, 400, 486, 988) comprises a fuel supply means for supplying fuel for delivery to the internal combustion engine, the fuel supply means having a fuel feed circuit (1262, 1322), that the fuel distribution means is in fluid communication with the storage means through the fuel feed circuit (1262, 1322) so that a sequential periodic fluid connection with the engine cylinders through the corresponding fuel injection lines (806, 1324, 1452, 1612, 1660) is possible, that the first electromagnetically actuated injection control valve (20, 440, 440, 454, 1406, 1604, 1634) in the fuel Delivery circuit (1262, 1322) is arranged between the storage means and the fuel distribution means for controlling the fuel injected into each engine cylinder during each of the sequential connection periods enabled by the fuel distribution means to thereby define sequential injection events, that the electromagnetically actuated injection control valve (20, 440, 440, 454, 1406, 1604, 1634) is movable between an open position in which a flow of fuel from the storage means to the fuel distribution means is possible and a closed position in which the flow of fuel from the storage means to the fuel distribution means is blocked, and that the pump arrangement (10, 400, 486, 988) comprises a quantity regulation control means arranged in the fuel delivery circuit (1262, 1322) between the storage means and the fuel distribution means to produce a predetermined, time-varying change in fuel pressure that occurs sequentially at each engine cylinder to perform an injection. 50. Pump arrangement according to claim 49, characterized in that the quantity regulation control means comprises a flow restriction means which is arranged in the fuel delivery circuit (1262, 1322) between the storage means and the fuel distribution means for limiting the flow of fuel from the storage means to the fuel distribution means during only a portion of each of the sequential injection events. 51. Pump arrangement according to claim 49 or 50, characterized in that the quantity regulation control means further comprises a side channel (1270, 1332, 1344, 1374) for directing the fuel flow around the flow restriction means and a quantity regulation circulation valve (1268, 1342) arranged in the side channel (1270, 1332, 1344, 1374), the quantity regulation circulation valve (1268, 1342) being movable into a closed position in which the fuel flow through the side channel (1270, 1332, 1344, 1374) is blocked and into an open position in which the flow through the side channel (1270, 1332, 1344, 1374) is possible. is. 52. Pump arrangement according to claim 51, characterized in that the flow restriction means has a fixed opening (1334, 1366) with a constant cross-sectional flow area for limiting the flow of fuel through the fuel delivery circuit (1262, 1322), or wherein the flow restriction means has a variable flow control valve (1266) which is movable between a first position in which a flow of fuel through the fuel delivery circuit (1262, 1322) is possible at a first flow rate and a second position in which a flow of fuel through the fuel delivery circuit is possible at a second flow rate. 53. A pump assembly according to claim 52, characterized in that the first flow rate occurs during a first portion of each injection event and the second flow rate occurs during a second portion of each injection event following the first portion, the first flow rate being greater than the second flow rate. 54. A pump assembly according to claim 51, 52 or 53, characterized in that movement of the flow control recirculation valve (1268, 1342) to the open position enables fuel to flow through the fuel delivery circuit (1262, 1322) at a third flow rate, the third flow rate being greater than the second flow rate, the third flow rate occurring during a third portion of each injection event following the second portion. 55. Pump assembly according to claim 52 or 53, characterized in that the variable flow control valve (1266) comprises a slidable piston (1272) having first and second ends (1280, 1282), a central bore (1284) having an inner end (1286) and an outer end, the outer end opening to the first end (1280) of the slidable piston (1272), the slidable piston (1272) having a plurality of openings (1294, 1296) extending from the inner end (1286) of the central bore (1284) through the second end (1282). 56. A pump assembly according to claim 55, characterized in that the variable flow control valve (1266) includes a tension spring (1300) operably connected to the slidable piston (1272) for biasing the slidable piston (1272) toward the first position, and preferably wherein the slidable piston (1272) is installed in a cavity arranged such that the slidable piston (1272) moves toward the second position whenever the upstream pressure exceeds the downstream pressure by a predetermined amount. 57. Pump arrangement according to one of claims 49 to 56, characterized in that the quantity regulation control means enables an increase in the fuel pressure in a corresponding fuel injection line (806, 1324, 1452, 1612, 1660) adjacent to the corresponding engine cylinder before and during each injection event at a first high rate followed by a low rate which is less than the first high rate followed by a second high rate, and preferably wherein the quantity regulation control means comprises a variable flow control valve (1266) which is movable between a first position in which the first high pressure rate is achieved and a second position in which the low pressure rate is achieved. 58. A pump assembly according to any one of claims 49 to 57, characterized in that fuel from the storage means can reach a maximum unrestricted flow rate corresponding to a maximum pressure in each of the fuel injection lines (806, 1324, 1452, 1612, 1660) adjacent to the corresponding engine cylinder during each injection event, the fuel delivery circuit (1262, 1322) having a first passage extending between the storage means and the injection control valve (20, 440, 440, 454, 1406, 1604, 1634), the injection rate control means comprising the first passage, the first passage having a predetermined length sufficient to provide a predetermined time delay between the movement of the electromagnetically actuated injection control valve (20, 440, 440, 454, 1406, 1604, 1634) into the open position and the achievement chen of the maximum pressure, wherein movement of the electromagnetically actuated injection control valve (20, 440, 440, 454, 1406, 1604, 1634) to the open position creates a pressure wave in the fuel delivery circuit (1262, 1322), the pressure wave traveling from the electromagnetically actuated injection control valve (20, 440, 440, 454, 1406, 1604, 1634) to an engine cylinder for defining or determining a wave motion time period, the predetermined length of the first passage being selected to obtain a desired wave motion time period. 59. Pump assembly according to claim 58, characterized in that the injection rate control means further comprises a second passage arranged parallel to the first passage for directing the flow from the accumulator means to the injection control valve (20, 440, 440, 454, 1406, 1604, 1634), and an orifice arranged in the second passage, and preferably wherein the quantity regulation control means allows an increase in fuel pressure in one of the fuel injection lines (806, 1324, 1452, 1612, 1660) adjacent to a corresponding engine cylinder during each injection event at a first high rate followed by a low rate that is less than the first high rate followed by a second high rate, the orifice providing an effective cross-sectional flow area for slowing the first high rate and the low Rate to desired values. 60. Pump arrangement according to one of the preceding claims, characterized in that the pump arrangement (10, 400, 486, 988) comprises a fuel supply means for supplying fuel for delivery to the internal combustion engine, the fuel supply means having a fuel feed circuit (1262, 1322), that the fuel distribution means is in fluid communication with the pump means through the fuel feed circuit (1262, 1322) to enable sequential periodic fluid communication with the engine cylinders through the respective fuel injection lines (806, 1324, 1452, 1612, 1660), that an injection control means is provided for controlling the fuel supplied to each engine cylinder during each of the sequential tial connection periods enabled by the fuel distribution means, thereby defining sequential injection events, and that a quantity regulation control means is provided which is arranged in the fuel delivery circuit (1262, 1322) between the pump means and the fuel distribution means for producing a predetermined, time-varying change in the rate of fuel injected into each engine cylinder during sequential injection events, the quantity regulation control means comprising a flow restriction means arranged in the fuel delivery circuit (1262, 1322) between the pump means and the fuel distribution means for limiting the flow of fuel from the pump means to the fuel distribution means during each sequential injection event, a bypass channel (1270, 1332, 1344, 1374) for directing the flow of fuel around the flow restriction means, and a quantity regulation bypass valve (1268, 1342) arranged in the side channel (1270, 1332, 1344, 1374). 61. Pump arrangement according to claim 60, characterized in that the quantity control circulation valve (1268, 1342) is movable into a closed position in which the fuel flow through the side channel (1270, 1332, 1344, 1374) is blocked and into an open position in which the fuel flow through the side channel (1270, 1332, 1344, 1374) is possible, and preferably wherein the flow restriction means comprises a fixed opening (1334, 1366) with a constant cross-sectional flow area for limiting the fuel flow through the fuel delivery circuit (1262, 1322), or wherein the flow restriction means comprises a variable flow control valve (1266) which is movable between a first position in which a fuel flow through the fuel delivery circuit (1262, 1322) with a first flow rate is possible, and a second position in which a fuel flow through the fuel delivery circuit (1262, 1322) with a second flow rate is possible. 62. Pump arrangement according to one of the preceding claims, characterized in that the pump arrangement (10, 400, 486, 988) a fuel supply means for supplying fuel for delivery to the internal combustion engine, the fuel supply means having a fuel delivery circuit (1262, 1322), the fuel distribution means being in fluid communication with the storage means through the fuel delivery circuit (1262, 1322) to enable sequential periodic fluid communication with the engine cylinders through the respective fuel injection lines (806, 1324, 1452, 1612, 1660), the electromagnetically actuated injection control valve (20, 440, 440, 454, 1406, 1604, 1634) being arranged in the fuel delivery circuit (1262, 1322) between the storage means and the fuel distribution means for controlling the fuel injected into each engine cylinder during each of the sequential communication periods defined by the fuel distribution means, and that a cavitation control means for minimizing cavitation in the fuel delivery circuit (1262, 1322) is provided between the cavitation control means and the cylinders, the cavitation control means comprising a backflow limiting valve (1460) arranged in the fuel delivery circuit (1262, 1322) between the injection control valve (20, 440, 440, 454, 1406, 1604, 1634) and the fuel distribution means to allow a substantially unobstructed forward flow of fuel to each engine cylinder while substantially limiting backflow. 63. Pump arrangement according to claim 62, characterized in that the pump arrangement (10, 400, 486, 988) further comprises an outlet passage for connection to the fuel delivery circuit (1262, 1322), wherein the electromagnetically actuated injection control valve (20, 440, 440, 454, 1406, 1604, 1634) is movable between an open position in which the flow of fuel from the storage means to the fuel distribution means is possible and a closed position in which the flow of fuel from the storage means is blocked, while the outlet passage is connected to the fuel delivery circuit (1262, 1322) downstream of the electromagnetically actuated injection control valve (20, 440, 440, 454, 1406, 1604, 1634) in fluid communication, wherein the backflow limiting valve (1460) is operable to provide a substantially unlimited flow of fuel from the electromagnetically actuated injection control valve (20, 440, 440, 454, 1406, 1604, 1634) to the fuel rail means when the solenoid-operated injection control valve (20, 440, 440, 454, 1406, 1604, 1634) is in the open position, and to limit the flow of fuel from the fuel rail to the solenoid-operated injection control valve (20, 440, 440, 454, 1406, 1604, 1634) when the solenoid-operated injection control valve (20, 440, 440, 454, 1406, 1604, 1634) is in the closed position. 64. Pump arrangement according to one of the preceding claims, characterized in that the pump arrangement (10, 400, 486, 988) comprises a fuel supply means for supplying fuel for delivery to the internal combustion engine, the fuel supply means having a fuel feed circuit (1262, 1322), that an outlet passage is provided for connection to the fuel feed circuit (1262, 1322), that the fuel distribution means is in fluid communication with the high pressure pump means through the fuel feed circuit (1262, 1322) so that a sequential periodic fluid connection with the engine cylinders through the corresponding fuel injection lines (806, 1324, 1452, 1612, 1660) is possible, that the injection control valve (20, 440, 440, 454, 1406, 1604, 1634) is arranged in the fuel delivery circuit (1262, 1322) between the high pressure pump means and the fuel distribution means for controlling the fuel injected into each engine cylinder during each of the sequential connection periods enabled by the fuel distribution means to thereby define sequential injection events, in that the injection control valve (20, 440, 440, 454, 1406, 1604, 1634) is movable between an open position in which fuel flow from the high pressure pump means to the fuel distribution means is possible and a closed position in which fuel flow from the high pressure pump means is blocked while the discharge passage is connected to the fuel delivery circuit (1262, 1322) downstream of the electromagnetically actuated injection control valve (20, 440, 440, 454, 1406, 1604, 1634) is in fluid communication, and that a void formation control means for minimizing void formation in the fuel delivery circuit (1262, 1322) and the fuel injection lines (806, 1324, 1452, 1612, 1660) are provided between the injection control valve (20, 440, 440, 454, 1406, 1604, 1634) and the engine cylinders, the cavitation control means being operable to maintain fuel in the fuel delivery circuit (1262, 1322) downstream of the fuel distribution means, the cavitation control means having an additional fuel supply (1600) connected to the outlet passage for supplying pressurized fuel at an additional supply pressure to the fuel delivery circuit (1262, 1322) downstream of the injection control valve (20, 440, 440, 454, 1406, 1604, 1634). when the injection control valve (20, 440, 440, 454, 1406, 1604, 1634) is in the closed position, wherein the auxiliary supply pressure is high enough to minimize the effects of cavitation while being sufficiently low not to cause fuel injection. 65. Pump arrangement according to claim 64, characterized in that the storage means is arranged along the fuel delivery circuit (1262, 1322) between the high pressure pump means and the injection control valve (20, 440, 440, 454, 1406, 1604, 1634) for collecting and temporarily storing fuel at high pressure received from the high pressure pump means. 66. A pump assembly according to claim 64 or 65, characterized in that the distributor housing (44, 410, 528, 713, 782, 1264, 1402) includes a rotor bore (214, 475), that the fuel distributor means includes a distributor rotor (216, 425, 462, 708, 1424) rotatably mounted in the rotor bore (214, 475), that the cavitation control means includes a refill means for refilling the plurality of injection lines (806, 1324, 1452, 1612, 1660), the refill means including a pressure booster pump means for supplying fuel at a booster pressure to the pump means, a pressure booster pump outlet passage fluidly connecting the pressure booster pump means to the pump means, and wherein a refill opening is formed in the distributor rotor (216, 425, 462, 708, 1424) and is in continuous fluid communication with the pressure booster pump outlet passage, wherein rotation of the distributor rotor tors (216, 425, 462, 708, 1424) causes the refill opening to periodically place the pressure boost pump outlet passage in fluid communication with each of the plurality of injection lines (806, 1324, 1452, 1612, 1660) to maintain the fuel in the plurality of injection lines (806, 1324, 1452, 1612, 1660) at the boost pressure. 67. Pump arrangement according to one of the preceding claims, characterized in that the pump arrangement (10, 400, 486, 988) comprises a fuel supply means for supplying fuel for delivery to the internal combustion engine, the fuel supply means having a fuel feed circuit (1262, 1322), that an outlet passage is provided for connection to the fuel feed circuit (1262, 1322), that the fuel distributor means is in fluid communication with the high pressure pump means through the fuel feed circuit (1262, 1322) so that a sequential periodic fluid connection with the engine cylinders through the corresponding fuel injection lines (806, 1324, 1452, 1612, 1660) is possible, that the injection control valve (20, 440, 440, 454, 1406, 1604, 1634) is arranged in the fuel delivery circuit (1262, 1322) between the high pressure pump means and the fuel distribution means for controlling the fuel injected into each engine cylinder during each of the sequential communication periods enabled by the fuel distribution means, the injection control valve (20, 440, 440, 454, 1406, 1604, 1634) being movable between an open position in which fuel flow to the fuel distribution means is possible and a closed position in which fuel flow from the storage means is blocked, while the discharge passage is in fluid communication with the fuel delivery circuit (1262, 1322) downstream of the injection control valve (20, 440, 440, 454, 1406, 1604, 1634) wherein the movement of the injection control valve (20, 440, 440, 454, 1406, 1604, 1634) from the open position to the closed position and from the closed position to the open position defines a run-out process and the movement of the injection control valve from the closed position to the open position and from the open position to the closed position defines an injection process, and that a cavity formation control means for minimizing the Cavitation is provided in the fuel delivery circuit (1262, 1322) and the fuel injection lines between the injection control valve and the engine cylinders, the cavitation control means comprising pressure regulating means provided in the spill passage for maintaining fuel in the fuel delivery circuit (1262, 1322) downstream of the injection control valve (20, 440, 440, 454, 1406, 1604, 1634) and in the fuel injection lines (806, 1324, 1452, 1612, 1660) at a regulated pressure during the spill process. 68. Pump assembly according to claim 67, characterized in that the storage means is arranged along the fuel delivery circuit (1262, 1322) between the high pressure pump means and the injection control valve (20, 440, 440, 454, 1406, 1604, 1634) for collecting and temporarily storing fuel at high pressure received from the high pressure pump means, preferably further comprising a refill passage (1678) in fluid communication with one end of the outlet passage between the injection control valve (20, 440, 440, 454, 1406, 1604, 1634) and the pressure regulating means and at an opposite end with the fuel distribution means, the fuel distribution means further serving to periodically connect the refill passage to the plurality of injection lines (806, 1324, 1452, 1612, 1660) in fluid communication so that fuel in the plurality of injection lines (806, 1324, 1452, 1612, 1660) is maintained at the regulated pressure, and wherein the pressure regulating means preferably comprises a cylinder (1636) having a first end and a second end, a piston (1640) slidably mounted in the cylinder (1636), and a biasing means for biasing the piston (1640) toward the first end so that fuel is forced into the refill passage (1678). 69. Pump arrangement according to one of the preceding claims, characterized in that the pump housing (22, 428, 500) contains an outwardly opening pump cavity (28, 30, 876, 878, 996), that a pump head (1170, 1189) is provided which can be fastened to the pump housing (22, 428, 500) around which outwardly opening opening pump cavity (28, 30, 876, 878, 996), wherein the pump head (1170, 1189) includes a pump unit recess (82, 84, 1116, 1118) arranged in connection with the pump cavity (28, 30, 876, 878, 996), and a valve cavity (74, 76, 908, 910, 1046, 1080) with a central axis aligned with the central axis of the pump unit recess (82, 84, 1116, 1118), that a pump unit (86, 88, 880, 882, 992, 993, 1112, 1114) is provided which is Pump unit recess (82, 84, 1116, 1118), the pump unit (86, 88, 880, 882, 992, 993, 1112, 1114) having a pump cylinder (116, 820, 844, 890, 892, 995) which includes the pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) and the pump plunger (122, 416, 418, 490, 492, 752, 884, 886, 1004) which is adapted to be reciprocally moved into the pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) and that a variable displacement control valve means is installed in the valve cavity (74, 76, 908, 910, 1046, 1080) for changing the effective displacement of the pump unit (86, 88, 880, 882, 992, 993, 1112, 1114) depending on a variable displacement control signal. 70. Pump arrangement according to one of the preceding claims, characterized in that the pump arrangement (10, 400, 486, 988) comprises a compact pump housing (22, 428, 500) with minimal dimensions in mutually perpendicular lateral, radial and axial directions, the pump housing (22, 428, 500) containing at least one pump cavity (28, 30, 876, 878, 996) with a first pump axis extending in the radial direction, that the pump means is installed in the pump cavity (28, 30, 876, 878, 996) for pressurizing fuel above the predetermined high pressure, that a pump head (1170, 1189) is provided which can be fastened to the pump housing (22, 428, 500) in order to pressurize the at least one Pump cavity (28, 30, 876, 878, 996), wherein the pump head (1170, 1189) comprises a bottom surface for engagement with the pump housing (22, 428, 500), an upper surface formed opposite the bottom surface, and a plurality of side surfaces connecting the bottom surface and the upper surfaces, that a pump control valve means (18, 19, 766, 994, 997, 1076, 1100, 1102) is attached to one of the plurality of side surfaces of the pump head (1170, 1189) for controlling the amount of fuel pumped into the at least one storage chamber (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232). 71. Pump assembly according to claim 70, characterized in that the pump assembly (10, 400, 486, 988) further comprises a pump valve recess (70, 72, 510, 512, 1104, 1106) formed in the one of the plurality of side surfaces and a pump valve cavity (74, 76, 908, 910, 1046, 1080) extending from the pump valve recess (70, 72, 510, 512, 1104, 1106) for receiving the pump control valve means, the pump control valve means comprising a reciprocating valve element which is arranged in the pump valve cavity (74, 76, 908, 910, 1046, 1080) is installed for movement along a valve axis which extends substantially transversely to the first pump axis, and preferably wherein the pump head (1170, 1189) and the storage means are formed in one piece or wherein the storage means has a storage housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) in which the at least one storage chamber (36a-g, 924, 926, 1012, 1134a-g, 1176, 1191, 1232) is formed, wherein the storage housing (34, 870, 990, 1078, 1110, 1174, 1187, 1230, 1244) is spaced from the pump head (1170, 1189) is arranged. 72. Pump assembly according to claim 71, characterized in that the pump assembly (10, 400, 486, 988) further comprises a pump valve cavity (74, 76, 908, 910, 1046, 1080) formed at least partially in the pump head (1170, 1189) for receiving the pump control valve means, the pump means comprising a pump cylinder (116, 820, 844, 890, 892, 995) having a pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) and a pump plunger (122, 416, 418, 490, 492, 752, 884, 886, 1004) mounted for reciprocating movement in the pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) along the first pump axis, the pump valve cavity (74, 76, 908, 910, 1046, 1080) is arranged radially between the pump chamber (124, 412, 414, 482, 484, 754, 828, 846, 872, 874, 1002) and the upper surface of the pump head (1170, 1189), wherein the first pump axis extends through the pump valve cavity (74, 76, 908, 910, 1046, 1080), and preferably wherein the pump cylinder (116, 820, 844, 890, 892, 995) is formed integrally with the pump head (1170, 1189) and/or wherein the pump control valve means comprises an annular valve seat which is attached to the pump cylinder (116, 820, 844, 890, 892, 995) is formed in the valve cavity (74, 76, 908, 910, 1046, 1080).