Conventional diesel fuel injection systems operate at injection pressures under 10,000 psi. For reasons arising primarily from the need to comply with increasingly more stringent engine emissions limits, efforts have been underway to develop injection systems that can operate at pressures above 20,000 psi. Many of these efforts are based on the so- called"common rail"configuration, wherein a pressurization subsystem produces and maintains a fuel pressure over 20,000 psi in an accumulator, and a delivery subsystem distributes and injects pressurized fuel from the accumulator to each engine cylinder.

A number of difficulties have thwarted the development of a commercially successful high pressure fuel injection system. A particularly troublesome difficulty has been the design of a rotary pump which can produce two to three times the pressure of a conventional rotary pump, without enlarging the overall exterior dimension, or "envelope", of a conventional pump.

A conventional rotary pump has plungers which reciprocate radially in corresponding pumping chambers. Fuel at inlet pressure is supplied through inlet passages to the pumping chambers, and fuel at outlet pressure is discharged through discharge passages from the pumping chambers. In the case of a pump which has radially outward pressurizing plungers, a rotating actuator periodically slides against the radially inner end of each plunger, periodically forcing the plunger and the fuel charge in the chamber, outwardly. The converse arrangement is present in a pump which has radially inward pressurizing plungers. The main design difficulties for achieving higher pressures, arise from the high torque loads and high friction generated by the sliding of the rotating actuator against

the plunger. Another difficulty is the loss of efficiency resulting from power wasted in moving some of the fuel through the passages and chamber without discharging such fuel at the outlet (i. e., the"dead volurne"problem).

The torque is a function of (a) the distance between the axis of rotation of the actuator and the sliding contact surface (i. e., friction radius), (b) the pressure in the pumping chamber, and (c) the cross sectional area of the pumping plunger. The lubrication requirements are dictated largely by the coefficient of friction, which depends strongly on variables (a) and (b). It can be readily appreciated that the outwardly pressurizing type of pump has an advantage relative to the inwardly pressurized pump, with respect to variable (a). However, in all outwardly pressurizing pumps known to the inventor, the fuel itself serves as the lubrication medium at the sliding surface. This presents a significant disadvantage, because fuel has a lower viscosity than engine lube oil, by at least a factor of ten. Thus, despite the short friction radius in conventional eccentric actuated, sequentially outward pumping multiple plunger pumps, the maximum achievable pressure is limited by the load capacity of fuel lubricated bushings, supporting the unbalanced pumping reaction force. However, such eccentric actuated pumps have the advantages of quiet operation and a low, uniform drive torque, resulting from long and overlapping pumping events.

It has thus been difficult to achieve the desirable combination of significantly higher pressure, while maintaining acceptable torque loads on the bearings and quiet operation. As the output pressure increases with a commensurate increase in torque load, while the size of the pump itself remains constant, or even decreases, relative to conventional pumps, it is not difficult to foresee the deleterious consequences arising from the combination of higher torque handled by a smaller bearing mounted in a relatively small pump body.

Summarv of the Invention It is therefore an object of the present invention to provide a rotary pump which can supply relatively high hydraulic pressure, with relatively low torque loads imposed on the pump.

It is a further object of the present invention to provide a rotary pump of such character, which can efficiently supply diesel fuel at pressures above 20,000 psi, within an exterior size no greater than conventional rotary pumps associated with diesel fuel injection systems.

It is yet another object of the present invention to provide a rotary pump of such character, which can be readily mounted on a vehicle drive train, to be driven by a rotating shaft of such drive train.

These and other objects and advantages are achieved in a vehicle having a drive train including an internal combustion engine with at least one shaft supported for rotation by internal bearing means, wherein the improvement comprises a high pressure hydraulic pump assembly rigidly connected to the drive train such that the drive member for the pump includes a cantilevered extension of the shaft from the bearing of the engine.

According to the invention, the pump uses bearings external to itself, from a power source which already exists in the conventional drive train, e. g., a cam shaft or crank shaft in the engine. In particular, the pump does not require its own load-carrying bearings. In the expected context in which the present invention is to be used, the driven shaft and associated bearings, are designed for extremely heavy loads, as compared to the loads generated by the pump. Therefore, the existing bearings for such shafts, can accommodate the additional loading resulting from operation of even a high pressure pump.

In the preferred context of a rotary fuel injection pump for a diesel engine, the pump can be mounted on the engine at the cantilevered end of the cam shaft, and the timing of the loads generated by the pumping action, can be coordinated with the loads imposed on the cam shaft during

engine valve operation, such that the maximum pump load is imposed on the bearings while the bearings are free from any load imposed by valve actuation.

This embodiment is more particularly embodied as a diesel fueled internal combustion engine having a rotatable cam shaft for controlling the operation of engine valves, wherein the cam shaft is supported in bearings within a shaft housing of the engine. A pump body is rigidly connected to the shaft housing. A plurality of pumping plungers are situated in a respective plurality of plunger bores oriented radially relative to a central axis through the body, thereby defining a pumping chamber in each bore. Plunger actuating means impart a periodically varying radial force on the plungers, whereby the plungers are reciprocated radially, thereby expanding and contracting each pumping chamber. A fuel supply is provided at relatively low pressure to the pumping chambers as they expand. Fuel discharge means deliver fuel at relatively high pressure from the pumping chambers as the chambers contract, for delivery under high pressure for injection into the engine. The cam shaft is cantilevered from the cam shaft bearing and coupled to the plunger actuating means, for imparting rotation thereto and thereby radially displacing the plungers.

The rotatable plunger actuating means, can be in the form of a circular ring eccentrically mounted relative to the central axis of the pump.

Alternatively, the plunger actuation means can be a cylinder projecting from the cam shaft, eccentrically relative to the central axis, or a coaxial projection that has a non-uniform exterior cam profile. In yet another implementation, the plunger actuation means can be a ring centered on the central axis, and with a non-uniform inner profile for imparting a radially inward force on the plungers.

Depending on the particulars of the pump design, the plungers could be actuated sequentially, or simultaneously. A pump having sequentially actuated plungers imposes unbalanced torque loads on the associated bearings, and therefore the present invention would have the

most advantageous use in connection with a rotary pump having sequentially actuated plungers.

In accordance with the preferred embodiment, wherein the pump is mounted on the engine valve housing, the further advantage of space minimization is achieved, for two reasons. First, the pump can be mounted in the housing, such that the pump is axially aligned with, and is substantially adjacent to, the free end of the cam shaft where supported in the cam shaft bearing. Secondly, the drive gear for the pump can be integrated with the drive gear for the cam shaft. As a result, the cam shaft for operation of the valves, and the drive gear for operation of the pump, are rotated simultaneously by a common belt, chain, or similar connection, as directly driven by the engine crank shaft. Thus, as a practical matter, the pump in accordance with the present invention, is directly driven by the engine and the loads generated by the pump are borne entirely by the engine.

It should be appreciated that by analogy, the pump according to the present invention, could be mounted elsewhere on or adjacent to the engine, while being driven by another rotating shaft powered by the engine. For example, the pump can be mounted to the engine block and penetrated by the forward extension of the crank shaft.

Another advantage of the present invention, arises from the intimate spatial relationship between the pump and the bearing of a driven shaft in the engine, such that the availability of lube oil for the bearing, provides a convenient source of lube oil for the main friction surfaces in the pump. Particularly with pumps of the type wherein the plungers are actuated radially inwardly, a reservoir and flow path for the lube oil of the engine, can be provided at the radially outer ends of the plungers, for lubricating the actuation surfaces.

Brief Description of the Draines The preferred embodiments of the invention will be described below with reference to the accompanying drawings, in which Fig. 1 is a schematic of an internal combustion engine, showing various locations at which a pump according to the invention can be mounted; Fig. 2 is a schematic illustrating in perspective, a portion of a pump mounted compact on an engine, for direct drive by the valve cam shaft; Fig. 3 is a longitudinal section view of a first embodiment of a pump mounted to the engine valve cover for direct drive by the valve cam shaft; Fig. 4 is a cross section view taken along line 44 of Fig. 3; Fig. 5 is an enlarged view of the pump valve arrangement in the embodiment shown in Fig. 3; Fig. 6 is a view similar to Fig. 4, showing the incorporation of conventional coil springs for biasing the cam shoes of the plungers, against the actuation ring; Fig. 7 is a view somewhat similar to Fig. 3, showing another embodiment of the pump, having an eccentric projection of the cam shaft, for driving the pump; Fig. 8 is a perspective view of the cam shaft and extension, in accordance with Fig. 7; and Fig. 9 is a schematic of another embodiment, wherein the pump is mounted on the engine for cooperation with the crank shaft.

Description of the Preferred Embodiments Figures 1 and 2 are schematic representations of an internal combustion engine 10, for example, as may be incorporated in a diesel powered vehicle. The engine crank shaft 20 delivers rotary power via an engine plate 22, to a transmission pressure plate (not shown), and via transmission gears, to the vehicle drive shaft. The crank shaft 20 is supported within the block portion 28 of engine 10, between crank shaft

bearings 30a, 30b. Pistons within cylinders bored in the engine block, in cooperation with fuel intake and combustion exhaust valves 14 provide the direct rotational torque on the crank shaft 20.

Above the engine block 28, an overhead valve housing or cover 32 surrounds a valve cam shaft 34, having lobes 36 which operate the engine cylinder valves, in a conventional manner. The cam shaft 34 is supported between bearings 38a, 38b. These bearings absorb torque loads imposed by the eccentric cam lobes 36 as they periodically bear against the resistance of the engine cylinder valves 36.

The cam shaft 34 is typically driven via a direct connection to the crank shaft drive gear 40, which projects from the crank case of the engine block 28 next to accessory pulley 42 for providing take off power to other rotating equipment. The cam shaft driven gear 44 is connected via a belt 50, chain or the like to the drive gear 40, such that the cam shaft 34 is rotated in synchronization with the crank shaft 20. The driven gear 44 is typically mounted within an end portion 16 or overhanging portion of valve cover 32.

In accordance with one embodiment of the present invention, a rotary hydraulic pump is directly connected to the valve cover 32 at locations 16 or 24, for direct drive by an axial extension of cam shaft 34.

As will be described in greater detail below, this arrangement avoids the necessity for the pump itself to be equipped with load carrying bearings associated with the pumping action; rather, these loads can be handled by the cam shaft bearing 38, externally of the pump. Another location where a pump can be mounted on the engine to take advantage of an engine bearing for the pump drive, is indicated at 18.

Figures 2-5 show the preferred embodiment of a high pressure pump 100 according to the present invention, mounted on the internal combustion diesel engine 10, for example at 16, as part of a common rail fuel injection system. In this embodiment, the pump 100 is rotatably driven directly by the cam shaft 34 which operates the intake and exhaust

valves 14 on the engine. A source of diesel fuel, such as a fuel pump from the fuel tank (not shown), supplies liquid fuel in the direction of arrow 116 at low pressure to the inlet 118 of the pump 100. The high pressure pump 100 delivers fuel at a pressure of at least about 20,000 psi in the direction of arrow 120, to the accumulator (not shown) of the common rail system. It should be understood, however, that the pump according to the invention can be connected to a different source of rotational drive, for delivery of a different kind of liquid at high pressure, for a different purpose.

The pump has a body 122 with an elongated hub portion 124 extending between arbitrary front and back ends 126,128 of the body.

The front of the body is preferably formed as flange or the like, for mounting to a rigid support structure such as the front 130 of the engine valve cover, which forms a cut out 48 defining a valve gear cover. The hub 124 has a central bore 132 extending from front to back, along a central axis 134 which in the mounted pump, is on an extension of the rotation axis of the engine cam shaft 34. The hub 124 has a plurality of plunger bores 136 spaced uniformly about the axis intermediate the front and back ends of the body, and extending radially through the hub portion to the central bore. The centerlines of the plunger bores 136 lie on a plane which, for convenience, will be referred to as the pumping plane 138. A fuel supply passage 140 extends obliquely from the front 126 of the body, through the hub portion 124, crossing from the front to the back of the pumping plane 138, and terminating at the central bore through optional further passage 142. Suitable fittings such as 118 can be provided at the front of the fuel supply passage, for connection to the low pressure fuel supply.

A valve housing 144 includes an elongated portion 146 situated in the central bore 132 of the body, in close coaxial relation within the hub portion 124, and a flange portion 148 in front of the hub portion, for rigidly engaging the flange portion 126 of the body 122, thereby fixing the valve

housing 144 both axially and angularly, relative to the body 122. A plurality of bolts 150, attach the flange portion 148 to the front of the body 122, for this purpose. The flanges on the body and valve housing permit assembly so that the various passages align axially and angularly, and load seals such as 184,186 to prevent leakage of fuel, especially at the front of the pump.

The valve housing hub 146 has a fuel inlet chamber 152 formed by an axial blind bore through the back end 154 of the housing, which is then plugged at 156 during fabrication of the pump. The fuel inlet chamber 152 is in fluid communication with the inlet passage, via a short inlet connecting passage 158 in the housing. The front end 160 of the inlet chamber 152, should be as close as possible to the pumping plane 138, for reasons explained more fully below. The valve housing has a discharge chamber 162 formed as an axial blind bore through the front end 148 of the housing. This is adapted to receive a suitable fitting 164 at the front end, for fluidly connecting the discharge chamber to, e. g., the accumulator of the common rail system. The back end or wall 166 of the discharge chamber, approaches the pumping plane 138.

The radially inner ends 168 of the plunger bores 136 are confronted by respective recesses 172 on outer surface 174 on the valve housing. The tolerances are maintained tight enough to establish a fluid seal between the bores 136 and the outer surface 174, such that the recess portion of the surfaces function as closure walls 176 for the bores 170. All the closure walls 176 are intercepted by the pumping plane 138.

The closure walls 176 can be shaped if desired, to enhance this sealing relationship. Because the fuel inlet chamber 152 is close to the pumping plane 138, a radius can be drawn from the central axis to the closure wall 176, through the inlet chamber 152. Short passages 170,184 are provided, to fluidly connect the inlet chamber 152 and the discharge chamber 162 to each plunger bore 136 at the closure wall 176.

A piston-like plunger 178 having radially inner and outer ends 180,182, is situated in each of the plunger bores 136, for reciprocal movement. The radial length of each bore will depend on the desired plunger stroke which, along with the bore diameter, defines the maximum volume of fuel which could be forced into the discharge chamber 162 at high pressure upon the plunger reaching its radially inner limit position.

Respective inlet check valve means fluidly connect the inlet chamber 152 with each of the plunger bores 136, through a respective closure wall 176, and respective outlet check valve means fluidly connect each plunger bore 136 with the discharge chamber 162, through the respective closure wall. The inlet check valve means includes a counter bored passage 170 defining an inlet port 190 which in part is fluidly connected to a plunger bore 136 through the closure wall 176 and in part covered by the hub 124, a valve seat 192 which tapers toward the fuel inlet chamber 152, and a ball element 194 situated in the counter bored passage. When the plunger 178 moves radially outward to draw fuel into the plunger bore 136, the ball element 194 moves radially outward into the contact with the hub 124 while maintaining the fluid connection between the inlet port 190 and the plunger bore 136. When the plunger moves radially inward to pressurize fuel in the plunger bore, the ball element 194 moves radially inward into contact with the valve seat 192 to prevent flow from the plunger bore 136 into the fuel inlet chamber 152.

The discharge check valve arrangement is situated in operative relation with each short discharge passage 184. The discharge chamber 162 has a back wall 166 which is perpendicular to the central axis 134, and a valve seat 196 is formed as a recess where each passage penetrates the back wall. A ball element 198 is sealable against a respective seat 196. Means are provided in the discharge chamber, for simultaneously biasing all the ball elements against their respective seats, to prevent opening as the inlet fuel fills the plunger bores. During actuation of the plungers in sequence, the balls 198 will be sequentially

forced out of their seats 196 by the very high pressure. Preferably, a piloted coil spring 200 is coaxially situated in the discharge chamber 162 to bear upon a flat disk 202 or the like, which in turn bears on all the ball elements 198. The disk can pivot slightly to accommodate the unseating of one ball, while maintaining the necessary seating force on the other balls. A stop 204 may optionally be provided for limiting the opening movement of the disk 202 It can be appreciated that the arrangement of the plunger bores 136, closure walls 176, inlet and discharge chambers 152,162, and associated connecting passages with valves, minimizes the dead volume of fuel which is subjected to the pressurization of the plungers, but which cannot be delivered to the discharge chamber. This advantage is achieved while permitting the inner limit position of the plungers during the pressurization stroke to closely approach the central axis 134. This helps minimize the torque radius, i. e., the distance from the central axis 134 to the actuation force applied at the radially outer ends 182 of the plungers 178.

At a radius intermediate the stroke outward limit position of the inner end 180 of the plungers 178 and the stroke inward limit position of the outer end 182 of the plungers, each plunger bore has a fuel leak off groove 206. These grooves drain away any fuel that might pass through the sealing effect of the tight tolerances between the plungers 178 and bores 136, and leading the fuel through the body to a leak off discharge port (not shown). Thus, all fuel in the pumping plane 138 is confined within a radius dictated by the leak off grooves 206. As a result, lube oil can be used to lubricate the plunger actuation surfaces.

The plungers 178 are actuated by a rigid actuating ring 208 which surrounds the plungers and is mounted for eccentric rotation about the central axis 134. The eccentricity drives each plunger inwardly in sequence, preferably via cam shoes 210 or the like, which facilitate the conversion of the rotary motion of the ring 208, into the linear motion of

the plungers 178. This conversion gives rise to a severe torque load, which tends to tilt the plunger axis relative to the bore axis, and generates an imbalanced force on the pump drive shaft 34 which rotates the actuating ring. The torque transmitted to the plungers and bearing 38a can be reduced by increasing the lubrication at the sliding contact surface 214 between the shoe and the actuating ring.

According to the embodiment shown in Figure 3, engine oil or other lube oil, which has a much greater viscosity than diesel fuel, can easily be provided to the sliding contact surface 214. The lube oil is supplied at the back end 128,154 of the body 122 and/or valve housing 144, or at the outer circumference of the actuating ring 210. The lube oil passes through the relatively wide axial tolerances or gaps 216, between the actuating ring and support structure 218 for the actuating ring. As can be appreciated upon inspection of Figure 4, lube oil can be supplied to the plunger actuating means, on the pumping plane and radially outside of the leak-off grooves. Openings 228 in the shoes provide tube oil to the captured end 182 of the plunger.

The support structure 218 preferably takes the form of (or can be integral with or in any event substantially surround) the driven or cam gear 44 that is already present for taking off power from the engine crank shaft 20 via drive gear 40, to rotate the valve cam shaft 34 (as shown in Figures 1 and 2). The external teeth of the support structure 218 are shown at 44 to engage a belt or chain 50 which in turn engages teeth on the gear 40 driven by the crank shaft. A circular collar 222 is rigidly connected via bolts 224 or the like, to the front face of the cam gear 218 in coaxial relation to the cam gear. The actuating ring 214 is rigidly mounted within the collar 222, eccentrically relative to the cam gear axis, so as to bear on the shoes 210.

With the cam shoes 210 in contact with the inner surface of the actuating ring 214 and the outer end 182 of each plunger 178, each plunger is driven to a radially inward limit position through a respective

plunger bore and thereafter each plunger must be permitted to move to a radially outward limit position, as the cam gear 218 is rotated. Fuel is thus periodically drawn at a relatively lower pressure from the inlet chamber 162 into each plunger bore 136 through a respective inlet check valve as each plunger moves toward its radially outer limit position and fuel is periodically delivered to the discharge chamber 162 at a relatively high pressure from each plunger bore through a respective discharge check valve as each plunger moves to its radially inner limit position. To assure that each plunger 178 moves to its radially outward limit position, energizer means can be provided, for biasing the plungers outwardly. In a typical arrangement wherein the outer end 182 of the plunger is captured to swivel within the shoe 210, the biasing means can act on the shoe.

Preferably, as shown in Figure 4, the energizer means is in the form of an elastic ring 226, pre-loaded compressively. The elastic energizing ring 226 circumscribes the valve housing portion 124 on the pumping plane 138 and maintains a radially outwardly directed bias against the inner sides of all the sliding shoes. The ring is preferably made from a material such as spring steel or Vespel (available from DuPont).

In the embodiment of Figure 3, the flange portion 126 of the body is rigidly mounted to the engine, at e. g., 130, providing the only support for the body 122 and pump valve housing 144 connected thereto. The axial position of the actuating ring 208, collar 222, and cam gear 218 are determined by the rigid engagement 232 of the cam gear 218 to the end 234 of cam shaft 34, which is cantilevered from one of the cam shaft bearings 38a. The bearing 38a is rigidly supported within the valve cover 32 or housing. The valve gear cover serves as a convenient mounting location for the body 122. The cam gear means 218 is thereby operatively connected to the body 122 and pump valve housing 144, only through the contact at 228 between the actuating ring 208 and the cam shoes 210.

Thus, the subassembly comprising body 122, pump valve housing 144, plungers 178, shoes 210, and shoe biasing means 226 are fixed axially independently of the axial fixing of the subassembly comprising the cam gear 218, collar 222, and actuating ring 208. In this manner, the space 216 or gap can be assured for providing paths for lube oil flow at the inner circumference of the actuating ring and the outer surface of the hub portion of the body, the latter flow helping to lubricate the radially outer portion of the plungers. Desired lube flow paths in the form of gaps on both axial sides of the actuating ring and shoes can be achieved by providing a smaller axial width for the actuating ring and shoes, than the axial width of the space between the body flange 126 and the cam gear 218.

Figure 2 shows the cam actuation gear 218 as the driven gear 44 which is concentrically connected to the extension 234 of the valve cam shaft within valve cover 32. The gear 218 as well as the cam shaft 34 are driven by the belt 50 and drive gear 40, which is rotated by the crank shaft 20 at the front of engine block 28. The pump has been omitted for clarity.

High pressure output of at least 20,000 psi can be achieved in a pump envelope which is no larger than, and can readily be made only about half as large as, conventional hydraulic supply pumps operating at about 7,000 psi discharge pressure. This is due to higher efficiency because of minimized dead volume and absence of wasteful nose volume spilling. A single fuel inlet chamber 152 and a single fuel discharge chamber 162 are connected with relatively short passages to the individual plunger bores 136. With these chambers and passages, and associated valves, all situated within a small diameter valve housing 144, i. e., radially inside of the inward limit position of the plungers 178, very little"dead space"arises. Moreover, these significant advantages are achieved in a configuration which provides smooth torque and quiet operation, due to the sequential actuation.

The minimization of the diameter of the eccentric 208 is facilitated in the preferred embodiment, by the circular energizing ring 226. The cross section of the body hub 124 is substantially circular, except for flattened regions 236 at the exterior, for the emergence of each plunger 178. Although three plungers are shown, a greater number, i. e., 6 or 8, can readily be achieved in accordance with the present specification. The width of the energizing ring 226 in the axial direction, is preferably approximately equal to that of the shoes 210. The energizing ring has holes or slots, which are penetrated by the outer ends 182 of the plungers, such that the plungers engage and capture the energizing ring, not unlike a sprocket engages mating holes on a tape or paper feed arrangement. The energizing ring 226 contacts all shoes 210 simultaneously. Therefore the dynamics of one shoe influences the dynamics of all other shoes, in a manner that requires a relatively small dynamic radius at the maximum outward position of the actuating shoe, relative to incorporation of a more conventional shoe energizing scheme.

Figure 6 shows an energizing arrangement 300, which is functionally similar to that of Figure 4, except that the energizing means for the cam shoes 302, incorporates conventional coil return springs 304.

Because these springs must be piloted, both the shoes and the hub 306, have projections 308,310 which extend radially toward each other. This increases the overall radius to the actuating ring 226', relative to the embodiment shown in Figure 4.

Figures 7 and 8 show an alternative embodiment, wherein the fundamental aspect of a main bearing external to the pump per se, is implemented in a pump arrangement 400 having radially outwardly actuated plungers. The pump 400 includes covers 402, surrounding a pump body 404. The body 404 has a front end 406 and a back end 408, the latter being attached, preferably with the pump housing 402, to the end portion 130 of engine valve cover 32. Within the pump body 404, a pumping assembly 408 is situated and supported such that a plurality of

radially extending plungers 412, are actuated by a rotating actuating means 414, in the form of an eccentric extension of the valve cam shaft 34. The actuator 414 sequentially urges each plunger outwardly, to compress the fluid in the plunger chamber for delivery at high pressure for e. g., injection into the engine or for some other purpose associated with the drive train or vehicle. The details of how the fluid, such as diesel fuel, is supplied to pump 400 and discharged from the pump, do not form a part of the present invention, but may be understood by reference to Figure 7 and an understanding of presently commercialized pumps similar in functionality to the pump described in, e. g., U. S. Patent No. 5,354,183.

As in the previously described embodiment, the plunger actuating means 414, is an extension of the portion 420 of the cam shaft 34, which is cantilevered from the cam shaft bearing 38. The cam shaft portion 420 is slipped through lip seal housing 418 mounted at the back end 408 of the pump body 404. If the pump 400 were mounted at location 24, shown in Figure 1, on the opposite side of the cam shaft 34, it would still be driven via gear 44. Thus, the belt 50 as shown in Figure 1 provides rotational drive to the cam lobes 36 on the one hand, and on the other hand to the plunger actuating member 414.

Figure 8 also shows that the plunger actuating means 414 in this embodiment, takes the form of a cylinder having a centerline 422 which is offset from the centerline 424 of the cam shaft 34.

It should be appreciated that a common feature of the two embodiments described above, is that the bearing for supporting the rotatable drive member for actuating the plungers, is external to the pump per se, and is moreover, a bearing which is normally provided for operation of the internal combustion engine.

Persons knowledgeable in the relevant technology, can readily understand from the foregoing description, that the pumps 100 (Figure 3) and 400 (Figure 7) can readily be mounted at location 24, as an alternative to location 16 (Figure 1). The location 18 associated with the

crank shaft 20 (Figure 1) is a desirable alternative, because of the greater inherent strength of the crank shaft and associated bearings 30, and because the crank shaft 20 typically rotates at twice the speed of the cam shaft 34.

Figure 9 shows a pump 500 mounted to the engine block 28 for direct actuation by the crank shaft 20 with pumping torque loads borne by the crank shaft bearing 30a. This corresponds to mounting location 18 of Figure 1. The crank shaft has a reduced diameter extension portion 502 on which is rigidly carried a cam ring 504 having an external profile 506 defining, for example, a plurality of convex lobes. The shaft extension 502 and cam ring 504 are concentric about the rotation axis 514 of the crank shaft 20. The various functional features of the pump 500 can take a variety of forms based on known pumps which are actuated by a centrally disposed, rotating cam such as 504. In particular, the generic pump 500 as illustrated, would be adapted for actuation of the pumping plungers, in a radially outward direction, e. g., with all or a plurality of the plungers actuating simultaneously for peak pressurization associated with the area of the cam at the crown of the lobes 506. Although a variety of implementing details are possible to one of ordinary skill in the art, the significant aspect of the present invention, is the mounting of the pump 500 between the engine crank case 28 and the accessories drive pulley 42 at the free end 516 of the crank shaft extension 502.

In the embodiment illustrated in Figure 9, lube oil is force fed through bearing 30a onto the external surface of the cam ring 504, to lubricate the contact surfaces with the associated plungers and or roller shoes (not shown). A leak off hole (not shown) is provided to maintain a steady lube oil quantity. The cam extension 502 can have a collar 508 located outwardly of the cam ring 504, to act as a rotating sealing surface against seal rings 510, which are seated at 512 in the pump 500. It should be appreciated that the lube oil is at relatively low pressure, and can therefore readily be confined within the radially inner portion of the

pump 500, whereas the high pressure fluid, such as fuel to be pumped to an accumulator for injection at, e. g., 20,000 psi, can be maintained in the radially outer portions of the pump 500, with dedicated high pressure check valve or other seals, as is known in the art.

As in the previously described embodiments, the plunger actuation is accomplished by rotation of the cam ring 504, which is rigidly supported and rotatably driven by the crank shaft extension 502, for imparting a periodically varying radial force on the plungers. The means 502,504,506 for actuating the plunger, may alternatively be considered as having a rotatable drive member 502,504 and a cam profile 506. The cam means can be integral with the crank shaft 20. Alternatively, the drive member can be integral with the crank shaft, and the cam profile rigidly connected to the drive member. As a further option, the drive member can be rigidly connected to the crank shaft, with the cam profile being either integral with or rigidly connected to the drive member. The crank shaft extension 502 is cantilevered from the bearings 30A, and the torque loads associated with the pump are borne by that bearing. The interaction of the seal 510 with the collar 508, does not perform a bearing function, and therefore does not transfer loads to the body of the pump 500.

The mounting of the pump 500 at the location shown in Figure 9, where the drive shaft 20,502 for the pump and the associated bearing 30a, are quite robust, in conjunction with the shaft rotation at engine speed rather than the one half engine speed of the traditional fuel injection pump, affords the designer the opportunity to make the pump 500 surprisingly compact. Such a compact, robust arrangement is particularly suitable for use in conjunction with a so-called common rail fuel injection system, wherein low pressure fuel is supplied to the pump 500 from a fuel supply tank 518, and the fuel is discharged from the pump 500 through a single discharge port and associated line 520, thereby delivering fuel at a relatively high pressure from the pumping chambers to the common rail fuel injection accumulator 522. Individual fuel injectors 524 are actuated

in a manner known in the art, to deliver particular quantities of fuel to the combustion chambers in the engine cylinder block, at or substantially near, the common rail pressure at 522. The pressurized fuel is discharged from pump 500 through line 520 to the accumulator 522, at a rate corresponding to the quantity of fuel withdrawn from the accumulator by the operation of the fuel injector.

It should be understood, however that although the embodiment shown in Figure 9 has only a single discharge port associated with the discharge line 520, a plurality of discharge ports could be provided, with an associated plurality of lines for delivering pressurized fuel to a plurality of accumulators, or directly to a plurality of fuel injectors.

It should also be understood that a pump 400 of the type shown in Figure 7 could be secured to the engine block 28 at location 18 as shown in Figure 1, and that it is not therefore necessary for the crank shaft extension to penetrate and pass through the pump according to the embodiment of Figure 9.