Radial piston pump with eccentrically driven roller actuating ring

Abstract

A high-pressure radial piston pump for fuel, comprising a hydraulic head (10) with two piston bores arranged symmetrically to a drive axis (16) and associated radial pistons (26a, 26b, 70, 80) and pump chambers (32), wherein the pistons (26a, 26b, 70, 80) and the pump chambers (32) are also symmetrically spaced apart by a cavity (12) arranged in the head (10), in which a rotatable eccentric drive element (18) with associated actuating ring (20) is arranged, wherein a rolling interaction is provided between the actuating ring (20) and the actuated ends (28, 82) of the pistons (26a, 26b, 70, 80) for intermittent actuation and a sliding interaction is provided between the actuating ring (20) and the drive element (18), wherein each piston bore has a center line (25a, 25b) includes, which intersects the actuating ring (20) but is offset by an offset (X) from the drive axis (16),viewed in cross-section perpendicular to the drive axis (16), characterized in that the piston bore center lines run parallel to each other, but are offset from each other by an offset (Y), viewed in a longitudinal section along the drive axis (16).

Description

[0001] The present invention relates to high-pressure radial piston pumps for fuels according to the preambles of patent claims 1 and 5.

[0002] Diesel common rail systems have become the state-of-the-art in the diesel engine industry and are currently entering their second and sometimes even third generations. Attention is currently focused on achieving further improvements in fuel economy and meeting more restrictive emissions regulations. In pursuing these goals, engine builders are more willing to select the most effective component for each part of the overall fuel injection system from a variety of suppliers rather than continuing to rely on a single system integrator.

[0003] DE 102 36 853 A1 discloses a high-pressure fuel pump having the features of the preambles of patent claims 1 and 5. The high-pressure fuel pump is designed with a plurality of pump pistons arranged at an angular distance from one another around a central drive shaft. The pump pistons, by means of preloaded springs, rest with their radially inner ends against the circumferential surface of the drive shaft and are each guided axially displaceably in a guide bore. With their end faces facing away from the drive shaft, the pump pistons each define a pump working chamber in the guide bore, which can be alternately connected to a fuel-filled low-pressure chamber and to a high-pressure delivery line. For improved high-pressure tightness, the guide bores of the pump pistons and the high-pressure channels are arranged directly in a common, one-piece pump housing.The guide bores thus form piston axes which are offset from the drive axis, and bow-shaped return means are provided for the pump pistons.

[0004] It is the object of the invention to improve a high-pressure fuel supply two- or three-radial piston pump in order to create a highly effective and universally applicable pump that can be installed in a wide variety of common rail injection systems.

[0005] According to the invention, this object is achieved as specified in claims 1 and 5, respectively. A hydraulic head comprises two or three radial pump pistons and associated pump chambers, which are arranged symmetrically opposite one another in the receiving body or annularly around a cavity in the head, in which an eccentric drive element with an associated outer rolling or rolling actuation ring is arranged, wherein a rolling or rolling interaction is provided between the actuation ring and the inner end of the piston for intermittent actuation, and a sliding interaction is provided between the actuation ring and the drive element. The corresponding inlet and exhaust valve controls are also arranged in the head, and the head is attached to an insert and / or a customer-specific mounting plate.

[0006] The drive element is rigidly supported by a drive shaft, which is supported by two bearings; one of the bearings is located in the mounting plate and the other in the hydraulic head. Depending on the current pumping force level and the rated speed, these bearings can be designed either as plain bearings or needle bearings. In the case of plain bearings, it is advantageous to make them fuel-lubricated by diverting a portion of the pressurized fuel from the feed circuit.

[0007] The actuating force for each pumping event is transferred from the eccentric to the pistons sequentially through the rolling actuating ring, which is supported by either a power-lubricated bearing or a needle bearing located approximately at the center of the shaft. The outer diameter of the rolling element is barrel-shaped to compensate for any offset or misalignment of the pistons relative to the drive shaft, for example, due to either tolerance stacking or bending.

[0008] Preferably, a semi-rigid yoke connects the pistons using a so-called desmodromic dynamic connection, forcing the inactive (non-pumping) piston toward bottom dead center while the other piston pumps. The rigidity of the yoke must be adjusted to minimize deflection (even at maximum vacuum during non-discharge conditions), as any separation and subsequent impact at the start of pumping has a detrimental effect on service life. At the same time, the contact force between the pistons and the outer diameter of the rolling element should be kept as low as possible to minimize wear and heat generation during intermittent sliding, which only occurs during the loading cycle.

[0009] In one embodiment, the pump has only two piston bores and associated two pistons, each piston bore has a centerline that intersects the actuating ring but is offset from the drive axis, and the piston bore centerlines are parallel to each other but offset from each other when viewed along the drive axis.

[0010] In another embodiment, the pump has three substantially equiangularly arranged piston bores and associated three pistons, and each piston bore has a centerline that intersects the actuating ring but is offset from the drive axis when viewed along the drive axis.

[0011] Preferably, each piston is arranged in its respective piston bore not only for free reciprocating movement along the bore axis during the charge and discharge phases of operation, but also for free rotation about the piston axis to counteract any asymmetric forces occurring at the interface between the radially inner end of the piston (or its associated shoe) and the actuating ring.

[0012] The pump output is preferably controlled by inlet metering with a proportional solenoid valve, but other conventionally available control methods may be used provided that the initial pressure of the inlet check valves is high enough to prevent uncontrolled and unwanted charging due to a vacuum created by the pistons during the vacuum stroke. To improve control resolution and thereby ensure complete control even at the lowest speeds, the control solenoid valve should be either a flow-proportional or pressure-proportional type combined with a variable curvature orifice.

[0013] The main advantages of the invention compared to currently available competitor pumps include: • Ability to generate high pump pressure up to 2000 bar. • Lack of a low-speed, high-force sliding interface between the piston and the rolling element. During partial discharge, which is the typical situation under normal operating conditions, mutual sliding occurs only during charging events and therefore only at a safe, low-force level. Even during the unusual operation in a 100% discharge mode (cold start), mutual sliding occurs at a reduced force level due to the unavoidable overlap of pressurized and depressurized strokes. • Lack of a preferential wear point at the drive shaft / rolling element, rolling element / piston, and piston / piston bore interfaces. During the pumping event, only rolling motion occurs between the piston and the rolling element. As pump output changes over time, the contact point changes, with statistically the entire inner and outer surface of the rolling element participating in the force transfer, resulting in a lower number of loading cycles at a given location. • Higher volumetric efficiency due to minimized occurrence of low pressure dead volume, reduced leakage due to maximized seal land lengths, lower number of leaks

[0014] interfaces and generally shorter pumping times, as well as increased pump chamber stiffness. • Higher mechanical efficiency. Low friction at the rolling interface combined with shorter piston overhangs, resulting in overall reduced friction losses. • Lower heat generation, resulting in reduced heat rejection (colder fuel). • Fewer parts and less complex machining, resulting in greater reliability and lower costs. Generally smaller and lighter pump. • Easier inlet metering control due to the absence of charge competition typical of pumps with overlapping charge events. • Reduced number of low and high pressure sealing interfaces. • Generally lower number of pumping cycles over the life of the pump. • Lack of return springs (a component subject to high dynamic stress) and lack of required installation space.

[0015] Further embodiments of the invention can be found in the following description and the subclaims.

[0016] The invention is explained in more detail below with reference to exemplary embodiments shown schematically in the figures. Fig. 1 is a schematic longitudinal sectional view of a two-piston pump according to the invention; Fig. 2 is a schematic cross-sectional view through the cavity of the Fig. 1 hydraulic head shown; Fig. Figure 3 is a graphical representation of the pump pressure versus the angle of drive shaft rotation obtained with the two-piston pump of Fig. 1; Fig. Figure 4 is a graphical representation of pump output versus drive shaft rotation angle for the pump of Fig. 1 at rated power and 1800 bar line pressure with inlet metering; Fig. Figure 5 is a longitudinal sectional view of the head of Fig. 1 with the additional features of a barrel-shaped actuating ring with the center of the apex in the same plane as the centerlines of the piston bores, viewed perpendicular to the drive axis; Fig. 6 is a view similar to Fig. 5, but with the center lines of the piston bores offset from the apex, viewed perpendicular to the drive axis; Fig. 7 is a cross-sectional view through the cavity of a hydraulic head for a three-piston pump design according to the invention; Fig. Figure 8 is a cross-sectional view through the hydraulic head of Fig. 7, including an early shut-off port with a check valve for each pump chamber; Fig. 9 is a schematic cross-section of a two-piston pump with a first alternative piston design; and Fig. Figure 10 is a schematic cross-section of a two-piston pump with a second alternative piston design.

[0017] Fig. 1 and Fig. 2 shows a high-pressure radial piston fuel pump having a hydraulic head 10 defining a central cavity 12 for receiving a rotatable drive shaft 14 disposed longitudinally along a drive axis 16 passing through the cavity 12. A cylindrical drive member 18 is rigidly supported by and spaced from the drive shaft 14 for eccentric rotation within the cavity 12 about the drive axis 16 as the drive shaft 14 rotates. A substantially cylindrical actuating ring 20 is annularly mounted around the drive member 18. A bearing 22, such as a needle bearing, is interposed between the drive member 18 and the actuating ring 20.

[0018] Two piston bores 24a, 24b extend in the head 10 to the cavity 12, each piston bore 24a, 24b having a centerline 25a, 25b intersecting the actuating ring 20 but offset by X from the drive axis 16 as viewed along the drive axis 16 (i.e., in a section perpendicular to the drive axis 16). A piston 26a, 26b is disposed in each piston bore 24a, 24b for free reciprocating movement and rotation therein. The pistons 26a, 26b have an actuated end 28 in the cavity 12 and a pumping end 30 remote from the cavity 12, the pumping end 30 cooperating with the piston bore 24a, 24b to form a pumping chamber 32. A piston shoe or shoe 34 extends rigidly from the actuated end 28 of each piston 26a, 26b and has an actuating surface for maintaining contact with the actuating ring 20 during rotation of the drive shaft 14.

[0019] Means are provided for biasing each piston 26a, 26b towards the cavity 12. This is preferably a semi-rigid yoke 36 arranged between the shoes 34 to dynamically coordinate (and thus ensure) the withdrawal of one piston 26a, 26b with the actuation of the other piston 26a, 26b according to a desmodromic effect. This also prevents backlash at low loads. The desmodromic yoke is not absolutely necessary for carrying out the general aspects of the invention, so that dedicated return springs for each piston 26a, 26b may be used (at additional cost and weight), or such biasing means may be omitted in some cases (as discussed below with respect to Fig. 10).

[0020] An inlet valve control 38 is provided in the head 10 for each pumping chamber 32 for delivering charge fuel through an inlet passage in the head 10 at a supply pressure to the pumping chamber 32. Similarly, an outlet valve control 40 is provided in the head 10 for each pumping chamber 32 for delivering pumped fuel to an outlet passage in the head 10 at a high pressure from the pumping chamber 32. Therefore, during one complete rotation of the drive shaft 14, each pumping chamber 32 undergoes two phases of operation. In a charge or intake phase, the corresponding piston 26a, 26b is retracted into the cavity 12 by the yoke 36, thereby increasing the volume of the pumping chamber 32 to accommodate an intake quantity of fuel from the inlet valve control 38.In the discharge or pumping phase, the corresponding piston 26a, 26b is actuated away from the cavity 12 by the actuating ring 20, whereby the volume of the pump chamber 32 is reduced and the amount of fuel for discharge is pressurized by the exhaust valve control 40.

[0021] The hydraulic head 10 includes a shaft mounting bore 42 for receiving one end 44 of the drive shaft 14, which is coaxial with the drive axis 16, and a bearing 46 for rotatably supporting this end 44 of the drive shaft 14. A removable mounting plate 48 is attached to the hydraulic head 10 and has a shaft mounting through bore 50 for receiving the other end 52 of the drive shaft 14, while the other end 52 is exposed for engagement with a source of rotational power. A suitable bearing 54 is provided in the mounting plate 48 for rotatably supporting the driven end 52 of the drive shaft 14. The mounting plate 48 may also have passages connected to the low-pressure supply pump for supplying a lubricating flow of fuel to the shaft bearings and the bearing between the eccentric drive member 18 and the actuating ring 20.

[0022] An important characteristic of the rolling relationship between the pistons 26a, 26b and the actuating ring 20 is that, although the actuating ring 20 always rotates (rolls) around the drive member 18 in the opposite direction of rotation of the drive shaft 14, this rotation is random, thereby preventing excessive wear at one location and also ensuring that lubricating fuel is quickly replenished at any location where metal-to-metal contact has occurred. Furthermore, the offset X of the piston bores 24a, 24b from the drive shaft axis reduces piston side loading.

[0023] Fig. Figure 3 is a graphical representation of the pump pressure as a function of the angle of rotation of the drive shaft corresponding to the two-piston pump of Fig. 1, which operates at a pressure of 1800 bar in the common rail and a pumping speed of 1000 min -1without intake metering. This represents a cold-start condition, which accounts for only a small fraction of the total time the engine is operating. The actuated ends 28 of the pistons 26a, 26b have a rolling interaction with the actuating ring 20, unless both pistons 26a, 26b pump simultaneously, as may occur briefly during cold start, at which point a sliding interaction occurs. Fig. 3 shows that over a narrow

[0024] An overlap pumping condition may exist within the angular range of drive shaft rotation (approximately 30 - 40 °), but the maximum pumping pressure during this overlap is less than 400 bar, which does not give rise to any worrying sliding friction.

[0025] Fig. Figure 4 is a graphical representation of pump output versus drive shaft rotation angle for the pump according to Fig. 1 at rated load and 1800 bar supply pressure with inlet metering. The strokes and displacements of the successive pistons 26a, 26b are indicated by C1 and C1', the regulated flow rate is indicated by C2, the average flow during pumping is indicated by C3, and the total average pumped flow is indicated by C4. This shows that the high pressure in each pumping chamber 32 is well separated during successive pumping events under rated load conditions.

[0026] Fig. Figure 5 shows a modification in which the actuating ring 20 has an outer surface 56 that has a roughly barrel-like shape. The curvature a rises and falls in the direction of the drive axis 16, and the apex 56' of the curvature α always remains in a plane defined by the center lines 25a, 25b of the two piston bores 24a, 24b.

[0027] The radius of curvature is very large, for example, on the order of one meter. Even with random or systematic deviations in the nominal parallelism between the centerline of the drive shaft 14 and the rotational axis of the actuating ring 20 and the nominal relationship between the piston centerlines and the rotational axis of the actuating ring 20 that arise during operation, the offset results in minimal piston side loading if the pumping force application point deviates only slightly due to eccentricity during pumping. However, this force application always occurs on the same section of the piston head. Therefore, the piston centerline is maintained in a coaxial relationship with the piston bore 24a, 24b.

[0028] Fig. Figure 6 shows two alternative designs. First, the centerlines of the piston bores 24a, 24b (although shown as coplanar, i.e., lying in the same plane) may instead be parallel but offset from each other, indicated generally by Y. Second, regardless of whether an offset Y is present, the apex 56" of the asymmetric radius of curvature (as shown) may lie in a plane parallel to, but offset by Z, from the centerlines 25a, 25b of the piston bores, viewed in longitudinal section perpendicular to the drive axis. This embodiment slightly increases piston side loadings, but it forces the piston 26a, 26b to rotate rather than slip during overlapping pumping events, thereby reducing the cumulative number of loading cycles at any given point on the shoes and actuating ring.

[0029] Fig. 7 and Fig. 8 show the invention in an embodiment in a three-piston pump with a drive shaft 16', piston bores 60a, 60b, 60c and pistons 62a, 62b, 62c. To avoid simultaneous pumping of two chambers, which would lead to high sliding forces at the shaft / piston head interface, a fixed pre-start orifice 66 delays the earliest start of pumping, resulting in separate pumping events. Essentially, the discharge phases of the pumping chambers subsequently occur as different pumping events, and each pumping chamber is fluidly connected to a pre-start orifice 66 for delaying the high-pressure fuel through the

[0030] The discharge passage associated with the pump chamber is ejected until the discharge of the high-pressure fuel through the discharge passage associated with the pump chamber of the previous pumping event is completed. Due to the shortened pumping duration for each of the three pumping events compared to only two pumping events, the discharge increases by only about 20% compared to the two-plunger pump with the same eccentricity and piston diameter, but the lower flow rate of the three pumping events per revolution reduces line pressure pulsation and provides greater flexibility in synchronizing the injection event and pumping event.

[0031] With the optional addition of a check valve 68 to the pre-pilot port 66, inlet metering output control can be performed through the same port 66. The check valve 68 in the pre-pilot port ensures separate pumping events and, at the same time, prevents backfilling due to a vacuum created by the retracted piston. Piston rotation induced by the off-center contact point is advantageous with and without pre-pilot because it continuously changes the contact point not only between the piston and the shaft, but also between the piston and its bore, thereby reducing the tendency for wear.

[0032] The three-piston pump may also include the design in which the center lines 64a, 64b, 64c of the piston bores 60a, 60b, 60c are offset from the axis 16', as shown in the cross section in Fig. 7, and the apex 56''' of the curvature in a plane parallel to, but offset by an offset Z' from, the centerlines 64a, 64b, 64c of the piston bores, as seen in the longitudinal section perpendicular to the drive axis. At the time when more than one piston is pumping (100% of maximum possible discharge), one or both pistons are allowed to rotate instead of slipping, thus protecting the piston-shaft interface from pre-ignition damage.

[0033] Fig. 9 shows alternative simplified pumping pistons or pistons 70 in piston bores 24, each piston 70 being an assembly comprising a rod 72 disposed in the piston bore with a shoe 74 integrally disposed therewith in the cavity and a substantially cylindrical sleeve 76 loosely surrounding the rod 72 and having a closed end 78 in the pumping chamber 32.

[0034] Fig.Figure 10 shows another piston embodiment, wherein each piston 80 is formed from a solid cylinder of a lightweight material, such as a ceramic, and has an actuated end 82 within the cavity 12 and a pumping end 84 remote from the cavity 12. The pumping end 84 cooperates with the piston bore 24 to define the pumping chamber 32, and the actuated end 82 maintains contact with the actuating ring 20 during rotation of the drive member 18. This embodiment can be operated without the biasing means because the vacuum associated with charging is sufficient to retract the piston during the charging phase of operation.

[0035] The same methods used for similar positive displacement pumps can be used to control the pump's output, such as inlet metering, pre-metering, early shut-off, post-shut-off, or a combination thereof.

Claims

[1] High-pressure radial piston pump for fuel, comprising a hydraulic head (10) with two piston bores arranged symmetrically to a drive axis (16) and associated radial pistons (26a, 26b, 70, 80) and pump chambers (32), wherein the pistons (26a, 26b, 70, 80) and the pump chambers (32) are also symmetrically spaced around a cavity (12) arranged in the head (10), in which a rotatable eccentric drive element (18) with associated actuating ring (20) is arranged, wherein a rolling interaction is provided between the actuating ring (20) and the actuated ends (28, 82) of the pistons (26a, 26b, 70, 80) for intermittent actuation and a sliding interaction is provided between the actuating ring (20) and the drive element (18), wherein each Piston bore comprises a center line (25a, 25b) which intersects the actuating ring (20) but is offset by an offset (X) from the drive axis (16),viewed in cross-section perpendicular to the drive axis (16), , characterized by that the piston bore center lines run parallel to each other, but are offset from each other by an offset (Y), viewed in a longitudinal section along the drive axis (16). [2] Pump according to claim 1, characterized by that corresponding inlet and outlet valve controls (38, 40) are also arranged in the head (10), the head (10) being attachable to a removable mounting plate (48) and the drive element (18) being rigidly carried by a drive shaft (14) which is supported by two bearings (54, 46), one of the bearings (54) being placed in the mounting plate (48) and the other bearing (46) in the hydraulic head (10). [3] Pump according to claim 1, characterized by that the outer surface (56) of the actuating ring (20) is barrel-shaped. [4] Pump according to claim 1, characterized bythat a semi-rigid yoke (36) connects the pistons (26a, 26b) and forces an inactive piston (26a, 26b) to the bottom dead center while the other piston (26a, 26b) pumps. [5] High-pressure radial piston pump for fuel, comprising a hydraulic head (10) with two or three individual radial pump pistons (26, 62, 72, 80) and associated pump chambers (32), wherein the pistons (26, 62, 72, 80) and the pump chambers (32) are annularly spaced around a cavity (12) arranged in the head (10) in which a rotatable eccentric drive element (18) with associated actuating ring (20) is arranged, wherein a rolling interaction is provided between the actuating ring (20) and the inner ends (28) of the pistons (26, 62, 72, 80) for intermittent actuation and a sliding interaction is provided between the actuating ring (20) and the drive element (18). wherein the hydraulic head (10) defines the central cavity (12) for receiving a rotatable drive shaft (14) arranged longitudinally along a drive axis (16) passing through the cavity (12); the drive element (18) is cylindrical and is rigidly supported offset from the drive shaft (14) for eccentric rotation in the cavity (12) about the drive axis (16) as the drive shaft (14) rotates; the actuating ring (20) is substantially cylindrical and annularly mounted around the drive element (18); Bearings (22) are arranged between the drive element (18) and the actuating ring (20), a piston (26a, 26b, 62) is arranged in each piston bore (24, 60) for free reciprocating movement therein, the piston (26a, 26b, 62) having the actuated end (28) in the cavity (12) and a pumping end (30) remote from the cavity (12), the pumping end (30) cooperating with the piston bore (24, 60) to define a pumping chamber (32); a piston shoe (34) extends rigidly from the actuated end (28) of each piston (26a, 26b, 62), the piston shoe (34) having an actuating surface for maintaining contact with the actuating ring (20) during rotation of the drive shaft (14); means (36) are provided for biasing each piston (26a, 26b, 62) toward the cavity (12); an inlet valve control (38) is provided for supplying charge fuel through an inlet passage in the head (10) at a feed pressure to the pump chamber (32); an exhaust valve control (40) is provided for supplying pumped fuel to a discharge passage in the head (10) at a high pressure from the pump chamber (32); wherein during a complete rotation of the drive shaft (14) each pump chamber (32) (a) a loading phase in which the associated piston (26a, 26b) is pulled towards the cavity (12) by the biasing means (36), thereby increasing the volume of the pump chamber (32) to accommodate an intake quantity of fuel from the intake valve control (38), and (b) an ejection phase in which the corresponding piston (26a, 26b) is actuated away from the cavity (12) by the actuating ring (20), whereby the volume of the pump chamber (32) is reduced and the amount of fuel for ejection is pressurized by the ejection valve control (40), and the piston bores (24, 60) extend in the housing through the cavity (12), and each piston bore (24, 60) has a center line (25a, 25b, 64a, 64b, 64c) which intersects the actuating ring (20) but has the offset (X) from the drive axis (16), viewed in a longitudinal section perpendicular to the drive axis (16), wherein the hydraulic head (10) has a shaft mounting bore (42) extending coaxially with the drive axis (16) for receiving one end (44) of the drive shaft (14) and a bearing (46) for rotatably supporting one end (44) of the drive shaft (14); and a removable mounting plate (48) is attached to the hydraulic head (10), the mounting plate (48) comprising a shaft mounting bore (50) for receiving the other end (52) of the drive shaft (14), while the other end (52) is subjected to engagement with a source of rotational energy, and bearings (54) for rotatably supporting the other end (52) of the drive shaft (14), characterized by that the actuating ring (20) has an outer surface (56) which is barrel-shaped, the curvature (α) rising and falling in the direction of the drive axis (16), wherein the apex (56') of the outer surface (56) lies in a plane defined by the center lines (25a, 25b, 64a, 64b, 64c) of the pumping bores, or wherein the apex (56") of the outer surface (56) lies in a plane parallel, but offset by an offset (Z), to the center lines (25a, 25b, 64a, 64b, 64c), viewed in a longitudinal view perpendicular to the drive axis (16). [6] Pump according to claim 5, characterized by that the pump has only three equally spaced piston bores (60a, 60b, 60c) and associated three pistons (62a, 62b, 62c), and each piston bore (60a, 60b, 60c) includes a center line (64a, 64b, 64c) which intersects the actuating ring (20) but is offset from the drive axis (16'), viewed in profile perpendicular to the drive axis (16'). [7] Pump according to claim 6, characterized bythat the discharge phase of the pumping chambers (32) occurs sequentially as a single pumping event and each pumping chamber (32) is fluidly connected to an early shut-off port (66) for delaying the discharge of high pressure fuel through the discharge passage associated with a given pumping chamber (32) until the discharge of the high pressure fuel through the discharge passage associated with the pumping chamber (32) of the previous pumping event is completed. [8] Pump according to claim 7, characterized by a check valve (68) in the opening (66) for early shut-off. [9] Pump according to claim 1 or 6, characterized by that the center lines (25a, 25b, 64a, 64b, 64c) are offset from one another when viewed in longitudinal section along the drive axis (16'). [10] Pump according to claim 1 or 6, characterized bythat the center of the apex lies in a plane defined by the center lines (25a, 25b) of the piston bores. [11] Pump according to claim 1 or 6, characterized by that the apex (56''') lies in a plane parallel to, but offset by an offset (Z') from, the center lines, viewed in longitudinal section along the drive axis (16'). [12] Pump according to claim 1 or 5, characterized by that each piston (70) is a composite comprising a rod (72) arranged in the piston bore (24) with a one-piece piston shoe (74) arranged in the cavity (12) and a substantially cylindrical sleeve (76) loosely surrounding the rod (72) and having a closed end (78) of the sleeve (76) at the pump chamber (32). [13] Pump according to claim 1, characterized bythat the hydraulic head (10) defines a central cavity (12) for receiving a rotatable drive shaft (14) arranged longitudinally along a drive axis (16) passing through the cavity (12); the drive element (18) is a cylinder rigidly supported by and displaced from the drive shaft (14) for eccentric rotation in the cavity (12) about the drive axis (16) when the drive shaft (14) rotates; the actuating ring (20) is substantially cylindrical and is annularly mounted around the drive element (18); Bearings (22) are arranged between the drive element (18) and the actuating ring (20), at least two piston bores (24a, 24b) extend in the hydraulic head to the cavity (12), each piston bore (24a, 24b) having a center line (25a, 25b) intersecting the actuating ring (20); a piston (26a, 26b) is arranged in each piston bore (24a, 24b) for free reciprocating movement therein, the piston (26a, 26b) having the actuated end (28) in the cavity (12) and a pumping end (30) remote from the cavity (12), the pumping end (30) cooperating with the piston bore (24a, 24b) to define a pumping chamber (32); a piston shoe (34) extends rigidly from the actuated end (28) of each piston (26a, 26b), and each piston shoe (34) has an actuating surface for maintaining contact with the actuating ring (20) during rotation of the drive shaft (14); means (36) are provided for biasing each piston (26a, 26b) toward the cavity (12); an inlet valve control (38) is provided for supplying charge fuel through an inlet passage in the head (10) at a feed pressure to the pump chamber (32); an exhaust valve control (40) is provided for delivering pumped fuel to a discharge passage in the head (10) at a high pressure from the pump chamber (32); wherein during a complete rotation of the drive shaft (14) each pump chamber (32) (a) a loading phase in which the associated piston (26a, 26b) is pulled towards the cavity (12) by the biasing means (36), thereby increasing the volume of the pump chamber (32) to make room for an intake quantity of fuel from the intake valve control (38), and (b) an ejection phase in which the associated piston (26a, 26b) is actuated away from the cavity (12) by the actuating ring (20), whereby the volume of the pump chamber (32) is reduced and the amount of fuel for ejection is pressurized by the exhaust valve control (40), undergoes, and the actuating ring (20) has an outer surface (56) which is barrel-shaped with a curvature (α) which rises and falls in the direction of the drive axis (16). [14] Pump according to claim 13, characterized by that the center (56') of the apex lies in a plane defined by the center lines of the pumping bores. [15] Pump according to claim 13, characterized by that the high point (56") of the apex extends in a plane parallel to, but offset from, the pump bore center lines, viewed in a longitudinal section perpendicular to the drive axis (16). [16] Pump according to one of claims 1-15, characterized bythat each piston (80) consists of a solid cylinder of a lightweight material, for example a ceramic, with an actuated end (82) in the cavity (12) and a pumping end (84) remote from the cavity (12), the pumping end (84) cooperating with the piston bore to define a pumping chamber (32), and the actuated end (82) maintaining contact with the actuating ring (20) during rotation of the drive shaft.

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