Abstract:
A high-pressure piston pump for a fuel system of an internal combustion engine, includes a housing, a piston, which defines a working chamber, and drive shaft having at least one crank section and supported in the housing by means of at least one shaft bearing. A piston bearing supports the piston at least indirectly against the crank section of the drive shaft. At least one of the bearings between parts that move in relation to one another is a hydrostatic bearing connected to the working chamber by means of a fluid connection. To increase efficiency, the fluid connection between the working chamber and the hydrostatic bearing is provided with a device operable to intermittently interrupt the fluid connection.

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
   This application is a 35 USC 371 application of PCT/DE 02/01888 filed on May 24, 2002. 

   BACKGROUND OF THE INVENTION 
   1. Field of the Invention 
   The current invention relates to a high-pressure piston pump for a fuel system of an internal combustion engine, with a housing, at least one piston that defines a working chamber, a drive shaft that is supported in the housing by at least one shaft bearing and has at least one crank section, and a piston bearing that supports the piston at least indirectly against the crank section of the drive shaft, wherein at least one of the bearings between parts that move in relation to one another is a hydrostatic bearing, which is connected to the working chamber by means of a fluid connection. 
   2. Description of the Prior Art 
   A pump piston of the type with which this invention is concerned, in the form of a radial piston pump, is known from DE 197 05 205 A1. In this radial piston pump, a bearing race is placed onto the eccentric section of a drive shaft. This bearing race has a flat contact surface against which a sliding block of an axially reciprocating piston rests. Between the contact surface of the bearing race and the sliding block, there is a relief chamber, which communicates with a working chamber defined by the piston via axial bores in the sliding block and in the piston. When the piston executes a delivery stroke, the pressure in the working chamber increases, which is conveyed through the bore in the piston to the relief chamber and thus leads to a reduction in the contact force between the sliding block and bearing race. The relief chamber thus constitutes a hydrostatic bearing. This reduces the friction and wear between the sliding block and bearing race. 
   Although the efficiency of the known piston pump during operation has in fact proven to be favorable, it is nevertheless not yet optimal. 
   The object of the current invention, therefore, is to modify a piston pump of the known type so that it has an even better efficiency. 
   This object is attained in a piston pump of the type mentioned above by virtue of the fact that the fluid connection between the working chamber and hydrostatic bearing is provided with a device that can intermittently interrupt the fluid connection. 
   SUMMARY OF THE INVENTION 
   The invention proceeds from the recognition that a leakage occurs in the vicinity of the chamber between the parts that move in relation to one another, i.e. fluid, which is to be supplied by the piston pump, travels as leakage fluid through the hydrostatic bearing and, for example, back to the inlet of the piston pump. This leakage is detrimental to the efficiency of the piston pump. It has also been established that it is not necessary to relieve the pressure on a bearing at all times during a work cycle of the piston pump. In essence, it makes sense to relieve the pressure of the bearing parts, which rest against each other and move in relation to each other, only at those times in which these two parts are pressed against each other with a relatively powerful force. In the case of a piston pump, this is essentially the case during the delivery stroke. 
   By providing the fluid connection between the working chamber and the hydrostatic bearing with a device that can intermittently interrupt the fluid connection, the invention makes it possible to sufficiently limit the time during which fluid flows from the working chamber into the hydrostatic bearing. This reduces the leakage quantity of fluid during operation of the piston pump without undesirably increasing the friction between parts of a piston pump bearing that move in relation to each other. Consequently, the efficiency of the piston pump is increased without shortening the service life of the piston pump. 
   The invention proposes including a pressure relief valve in the device that can intermittently interrupt the fluid connection. This pressure relief valve is incorporated into the fluid connection so that it opens this fluid connection only if the pressure in the region of the fluid connection oriented toward the working chamber exceeds a threshold value. This is based on the concept that the stresses on the bearings are at their greatest when the pressure in the working chamber is high. A piston pump of this kind is simple in design and operates reliably. 
   It is also possible to include an on-off valve in the device that can intermittently interrupt the fluid connection. In this modification, therefore, it is possible to select at will the times at which the hydrostatic bearing is connected to the working chamber and the times at which this connection is interrupted. This permits the fluid quantity used for the hydrostatic bearing to be reduced even further. 
   In this connection, it is particularly preferable if the on-off valve is the quantity control valve of the piston pump. A quantity control valve of this kind is usually used to temporarily short-circuit the outlet of the piston pump to its inlet toward the end of a delivery stroke, thus limiting the quantity of the effectively delivered fluid. In this modification, hardly any fluid is lost to produce the hydrostatic bearing since the production of this hydrostatic bearing uses only the fluid, which, in order to limit the delivery quantity, is not supposed to travel to the actual outlet of the piston pump anyway, but is conveyed back to its inlet. 
   The piston pump according to the invention is relatively small if the device that can intermittently interrupt the fluid connection is accommodated in the piston. However, it is also possible to accommodate it in the housing of the piston pump. This makes it easier to access the device, e.g. for maintenance purposes. 
   The considerably reduced fluid quantity required to generate a hydrostatic bearing in the piston pump according to the invention makes it possible to embody several or possibly even all of the highly stressed bearings in the piston pump with such a hydrostatic bearing. This potential is realized by the modification in which at least one hydrostatic bearing is respectively provided in the piston bearing and in the shaft bearing. 
   The hydrostatic bearing can contain a chamber, which is limited in the azimuth direction. This reduces the volume of the chamber and consequently reduces the fluid quantity required to generate a hydrostatic bearing. Such a limitation of the chamber does not result in any significant increase in the bearing friction forces since the hydrostatic bearing only has to work in the direction of the force peaks. These peaks naturally occur primarily when the piston is disposed in the vicinity of its top dead center and the fluid enclosed in the working chamber is thus maximally compressed. 
   The piston pump according to the invention can be embodied as a single cylinder piston pump and as a multicylinder piston pump. The angular range over which the chamber extends in the azimuth direction is preferably less than 360°/2 times the number of pistons. 
   The length and the width of the chamber are used to produce a hydrostatic bearing that is optimal for each individual application. 
   Another modification is characterized in that the fluid connection is connected to a pressure damper. This pressure damper can be embodied as a compression volume, spring bellows, diaphragm chamber, or the like. Such a pressure damper can be used to shape the chronological course of the fluid flow that flows from the working chamber to the chamber. This is particularly advantageous if the device that can intermittently interrupt the fluid connection is the quantity control valve of the piston pump. If this quantity control valve is opened toward the end of the delivery stroke, then an abrupt pressure increase occurs in the fluid connection and consequently also in the chamber. This pressure increase can be flattened somewhat by means of such a pressure damper. 
   This goal is shared by the modification in which at least one flow throttle is provided between the fluid connection and the pressure damper. For example, when a pressure relief valve or an on-off valve is used, such a flow throttle reduces the chronological pressure gradient in the fluid connection and extends the time of the pressure increase somewhat. The hydrostatic bearing is consequently available for a longer time than the fluid connection is open between the chamber and the working chamber. 
   The fluid connection to the chamber in the shaft bearing can include a flow conduit in the housing, which is connected to an annular groove in a bearing shell or in the shaft, which annular groove is connected to a radial bore in the shaft, which radial bore is connected to an axial bore in the shaft, which axial bore is connected to a radial bore in the shaft, which radial bore feeds into the chamber in the shaft bearing. Bores of this kind are easy to produce, which simplifies the production of the fluid connection. 
   The same is also true for the fluid connection, which leads to the chamber in the piston bearing and which includes a radial bore that leads away from the axial bore in the shaft and feeds into the chamber in the piston bearing. 
   The invention also relates to a fuel system for an internal combustion engine, with a fuel tank, a fuel pump that feeds into a fuel accumulation line, and at least one fuel injection device that is connected to the fuel accumulation line and injects the fuel directly into the combustion chamber of an engine. 
   In order to increase the efficiency of such a fuel system, the invention proposes that the fuel pump be embodied in the above-described manner. 
   The invention also relates to an internal combustion engine with at least one combustion chamber into which the fuel is directly injected. Such an engine is advantageously provided with a fuel system of the type mentioned above. 

   
     BRIEF DESCRIPTION OF THE DRAWINGS 
     Exemplary embodiments of the invention will be explained in detail below in conjunction with the accompanying drawings, in which: 
       FIG. 1  is a schematic representation of a fuel system with a first exemplary embodiment of a fuel pump according to the invention; 
       FIG. 2  is a partially sectional representation of the fuel pump from  FIG. 1 ; 
       FIG. 3  shows a section along the line III—III from  FIG. 2 ; 
       FIG. 4  shows a section along the line IV—IV from  FIG. 2 ; 
       FIG. 5  is a representation of the angular range of a force vector of the fuel pump from  FIG. 2  in relation to the longitudinal axis of a drive shaft; 
       FIG. 6  is a representation similar to  FIG. 1  of a fuel system with a second exemplary embodiment of a fuel pump; 
       FIG. 7  is a representation similar to  FIG. 2  of the fuel pump from  FIG. 6 ; 
       FIG. 8  is a representation similar to  FIG. 1  of a fuel system with a third exemplary embodiment of a fuel pump; 
       FIG. 9  is a representation analogous to  FIG. 3  of the corresponding region of the fuel pump from  FIG. 8 ; 
       FIG. 10  is a representation analogous to  FIG. 4  of the corresponding region of the fuel pump from  FIG. 8 ; and 
       FIG. 11  is a representation of the angular range of a force vector of the fuel pump from  FIG. 8  in relation to the longitudinal axis of a drive shaft. 
   

   DESCRIPTION OF THE PREFERRED EMBODIMENTS 
   In  FIG. 1 , a fuel system is labeled as a whole with the reference numeral  10 . It is part of an internal combustion engine  11  and includes a fuel tank  12  from which an electric fuel pump  14  delivers the fuel into a fuel line  16 . This fuel line  16  leads to an inlet  18  of a high-pressure fuel pump, which is labeled as a whole with the reference numeral  20  and which is driven by a crankshaft, not shown, of the internal combustion engine  11 . The precise design of this high-pressure fuel pump will be discussed in detail below. 
   From an outlet  22 , a fuel line (no reference numeral) leads to a fuel accumulation line  24 , which is commonly also referred to as a “rail”. A number of fuel injection devices  26  are connected to the fuel accumulation line  24 . These devices are high-pressure injection valves or injectors. The latter are connected to the engine block (not shown) of an internal combustion engine (not shown) and inject the fuel directly into combustion chambers  28 . 
   A pressure sensor  30  detects the pressure in the fuel accumulation line  24  and sends a corresponding signal to a control and regulation unit  32 . In a manner that is not shown in detail, this unit in turn is connected at its output end to the high-pressure fuel pump  20 . The high-pressure fuel pump  20  is a radial piston pump with three cylinders arranged in a star pattern. In principle, the high-pressure fuel pump  20  is designed as follows: 
   From the inlet  18 , a flow conduit  34  leads through a check valve  36  to a branch point  38 . The check valve  36  opens inward and thus protects the fuel line  16  and the electric fuel pump  14  from pressure surges. From the branch point  38 , flow conduits lead to the individual cylinders  40   a ,  40   b , and  40   c . The cylinders  40   a - 40   c  are identically designed. For the sake of clarity, reference numerals are furnished for only one of the cylinders. 
   Each cylinder  40   a - 40   c  has a check valve  42  on the inlet side, a pump unit  44 , and a check valve  46  downstream of the pump unit  44 . Downstream of the check valves  46 , the flow conduits of the individual cylinders  40   a - 40   c  come back together at a junction point  48 . From there, a flow conduit  50  leads through another check valve  52  to the outlet  22  of the high-pressure fuel pump  20 . 
   A flow conduit  54  branches off from the flow conduit  50  between the junction point  48  and the check valve  52  and this flow conduit  54  contains an on-off valve  56 . This on-off valve is an electrically actuated 2/2-way on-off valve, which is open in its neutral position  58  and is closed in its actuated position  60 . The control and regulation unit  32  controls the on-off valve  56 . The flow conduit  54  leads from the on-off valve  56  to a hydrostatic bearing  62 , which will be explained in detail below. 
   A flow conduit  64  branches off from the flow conduit  54  downstream of the on-off valve  56  and at its other end, this flow conduit  64  feeds into the flow conduit  34 , between the check valve  36  and the branch point  38 . The flow conduit  64  contains a pressure damper  66 , which in this instance is a spring/piston chamber. However, it is also possible to embody the pressure damper  66  as a compression volume, spring bellows, diaphragm chamber, or the like. A first flow throttle  68  is provided upstream of the pressure damper  66  in the flow conduit  64  and another flow throttle  70  is provided downstream of the pressure damper  66  in the flow conduit  64 . 
   The precise embodiment of the high-pressure fuel pump  20  can be inferred from  FIGS. 2-4 . It should be noted that only one cylinder  40  is depicted in this intersecting plane and that individual conduits, etc. are not visible. 
   The high-pressure fuel pump  20  has a housing  72 . This housing contains a blind bore-like recess  74  whose longitudinal axis extends horizontally in FIG.  2 . The housing  72  also contains another recess  76 , which extends vertically in  FIG. 2 , from the upper edge of the housing  72  into the horizontal recess  74 . The horizontal recess  74  contains a drive shaft  78 . This shaft is connected to the crankshaft (not shown) of the internal combustion engine. 
   The drive shaft  78  is supported in the vicinity of each of its two longitudinal ends by a bearing in the housing  72 . The bearing on the left in  FIG. 2  is labeled with the reference numeral  80 . To the right of the bearing  80  in  FIG. 2 , the horizontal recess  74  is sealed in relation to the outside by a shaft seal  82 . The right end of the drive shaft  78  is supported in a hollow, cylindrical bearing shell  84 , which constitutes a shaft bearing. Approximately in its middle in the axial direction, the drive shaft  78  has an eccentric section  86 , which is placed against a bearing race  88 . 
   The vertical recess  76  is closed at the top by a cover  90 . A guide sleeve  92  is inserted into the recess  76 . This guide sleeve  92  in turn guides a piston  94  in an axially movable fashion. A foot  96  is welded to the bottom end of the piston  94  in  FIG. 2. A  compression spring  98  is clamped between the foot  96  and guide sleeve  92 . This spring presses the foot  96  and consequently also the piston  94  against the bearing race  88 . The bearing race  88  consequently constitutes a piston bearing (no reference numeral) that supports the piston  94  in relation to the drive shaft  78 . 
   A working chamber  100  is provided above the piston  94  in FIG.  2 . This chamber is fed from the left in  FIG. 2  by the flow conduit that contains the check valve  42 . The flow conduit that contains the check valve  46  extends from the working chamber  100  toward the right in FIG.  2 . Neither the branch point  38  nor the junction point  48  is visible in the intersecting plane depicted in FIG.  2 . The working chamber  100  and the piston  94  are part of the pump unit  44  of the cylinder  40  depicted. 
   The hydrostatic bearing  62  is designed as follows: 
   From the on-off valve  56 , the flow conduit  54  leads to the horizontal recess  74 . By means of a bore  102  in the bearing shell  84 , the flow conduit  54  continues to an annular groove  104  on the inside of the bearing shell  84 . At the same axial position as the annular groove  104 , a radial bore  106  is let into the drive shaft  78  and feeds into an axial bore  108  in the drive shaft  78 . This axial bore  108  extends into the eccentric section  86  of the drive shaft  78 . 
   A radial bore  110  leads outward from the axial bore  108  to a recess (no reference numeral) on the outer circumferential surface of the drive shaft  78 . As can be seen in  FIG. 3 , this recess extends in the azimuth direction over an angular range of approximately 60° (for the sake of clarity, only the shaft  78  and the bearing shell  84  are shown in  FIG. 3 ; in an exemplary embodiment that is not shown, the angle is less than 60°). This produces a chamber  112  in which a hydrostatic counteracting force, which counteracts the forces coming from the piston  94 , is generated in a manner that will be explained below. 
   In the same manner, but offset by 180°, a radial bore  114  branches outward from the axial bore  108  in the vicinity of the eccentric section  86 , and in an analogous manner, feeds into a chamber  116 . As shown in  FIG. 4 , this chamber  116  also extends in the azimuth direction over an angular range of approximately 60° (in an exemplary embodiment that is not shown, this angle is less than 60°). Here, too,  FIG. 4  depicts only the shaft  78  and the bearing race  88  for the sake of clarity. 
   The high-pressure fuel pump  20  functions as follows: 
   Because of the eccentric section  86 , a rotation of the drive shaft  78  sets the piston  94  into an axial reciprocating motion. The control and regulation unit  32  triggers the on-off valve  56  so that it is closed at first during a delivery stroke of the piston  94 , i.e. when the piston is moving upward. This increases the pressure of the fluid enclosed in the working chamber  100  considerably. By means of the flow conduit  50 , which is not visible in  FIG. 2 , the compressed fluid travels out of the working chamber  100  into the fuel accumulation line  24 . The pressure sensor  30  detects when the desired pressure in the fuel accumulation line  24  has been achieved. 
   The control and regulation unit  32  then triggers the on-off valve  56  so that it opens. As a result, the fluid connection opens between the working chamber  100  and the chambers  112  and  116  of the hydrostatic bearing  62 . This increases the pressure in the chambers  112  and  116 , which generates a hydrostatic counteracting force in the desired direction between the bearing shell  84  and the drive shaft  78  (shaft bearing) and on the other hand between the bearing race  88  and the drive shaft  78  (piston bearing). At the end of the delivery stroke, the control and regulation unit  32  closes the on-off valve  56  again, which interrupts the fluid connection once more between the working chamber  100  and the two chambers  112  and  116 . 
   However, the closing of the on-off valve  56  does not immediately terminate the hydrostatic counteracting force generated in the chambers  112  and  116 . First of all, it takes a certain amount time for the fluid to drain out through the gaps on the one hand between the drive shaft  78  and the bearing shell  84  and on the other hand between the drive shaft  78  and the bearing race  88 . Secondly, the pressure damper  66  functions as a pressure reservoir, which continues to supply a certain quantity of fluid into the chambers  112  and  116  even when the on-off valve  56  is closed. 
   The chronological progression of the hydrostatic counteracting force generated by the pressure buildup in the chambers  112  and  116  is determined on the one hand by the width and the azimuth angular span of the chambers  112  and  116  and on the other hand by the properties of the pressure damper  66  and the two flow throttles  68  and  70 . As mentioned above, the azimuth angular span of the chambers  112  and  116  is maximally 60°; in any case in a multicylinder pump, this angular span is maximally 360°/2 times the number of cylinders, or 60° with the three cylinders here. This angular span is a result of the following considerations: 
   As shown in  FIG. 5 , the force vector resulting from the exertion of pressure on the pistons of the cylinders  40   a  to  40   c  in the current three-cylinder high-pressure pump  20  varies in a range of approximately 60° depending on the angular position of the drive shaft  78 . The beginning of the range is once again offset by approximately 60° in the rotation direction (arrow  121  in  FIGS. 4 and 5 ) in relation to an axis  122 , which rotates with the shaft and points in the eccentricity direction. Within the above-mentioned angular range, the force vector rotates synchronously with the drive shaft  78  around its longitudinal axis. Starting from this loading phase, the unloading phase occurs by means of the hydrostatic force on the piston bearing (bearing race  88  and shaft  78 ) in the vicinity of the chamber  116  and on the shaft bearing (bearing shell  84  and shaft  78 ) offset from this by 180°, in the vicinity of the chamber  112 . 
   In the exemplary embodiment shown in  FIGS. 1  to  5 , the hydrostatic bearing  62  has hardly any negative influence on the efficiency of the pump  10  since the hydrostatic bearing  62  is produced using only fluid, which the on-off valve  56  is already expending anyway for pressure control. Therefore no additional leakage is required to produce the hydrostatic bearing. 
     FIGS. 6 and 7  show a second exemplary embodiment of a high-pressure fuel pump  20 . Parts, elements, and regions, which have functions equivalent to those of parts, elements, and regions described previously, have been provided with the same reference numerals and are not explained again in detail. 
   By contrast to the exemplary embodiment described above, instead of an on-off valve, a pressure relief valve  118  is disposed in the fluid connection  54  between the working chamber  100  and chambers  112  and  116 . This pressure relief valve  118  opens the fluid connection  54  only when the pressure in the working chamber  100  exceeds a certain threshold value. As a result, the hydrostatic counteracting force only becomes fully effective above the opening pressure of the pressure relief valve  118 . 
   The advantage to this is that—without the need for an electric triggering—at low pressures in the working chamber  100 , no fluid flows in the form of leakage through the chambers  112  and  116  and the corresponding bearing gaps on the one hand between the drive shaft  78  and the bearing shell  84  and on the other hand between the drive shaft  78  and the bearing race  88 , which results in a higher volumetric efficiency of the high-pressure fuel pump  20 . In the upper pressure range, a higher leakage does in fact occur, but this is at least compensated for with regard to the overall efficiency due to the lower bearing load and the resulting higher mechanical efficiency. In any case, independent of the efficiency, this results in a considerably extended service life of the high-pressure fuel pump  20 . 
   In addition to the first exemplary embodiment, an additional axially extending groove  120  is provided on the inside of the bearing shell  84 . This groove extends from the chamber provided to the right of the bearing shell  84  to the space in the recess  74  provided to the left of the bearing shell  84 . The groove  120  prevents a pressure buildup from occurring at the end face due to the leakage between the drive shaft  78  and the bearing shell  84 , which could produce impermissibly high axial forces on the drive shaft  78 . The space provided in the horizontal recess  74  to the left of the bearing shell  84  is connected in a manner not shown in detail here to the inlet  18  of the high-pressure fuel pump  20 . 
     FIG. 8  shows another exemplary embodiment of a high-pressure fuel pump. Here, too, components and regions whose functions are equivalent to those of corresponding components and regions in the preceding figures are provided with the same reference numerals and are not explained again in detail. 
   In contrast to the exemplary embodiments shown in  FIGS. 1 and 6 ,  FIG. 8  depicts a 1-cylinder piston pump  20 . Among other things, this also results in a different orientation of the chambers  112  and  116 , as shown in  FIGS. 9 and 10 . According to them, the chamber  116  is disposed in a range of approximately 60° on both sides of the eccentricity axis  122 . It therefore has approximately twice the angular span of the corresponding chamber in the preceding exemplary embodiments. In addition, it is offset by 90° counter to the rotation direction of the drive shaft  78  in comparison to the preceding exemplary embodiments. The chamber  112  is offset from the chamber  116  by 180°, i.e. is disposed with its center axis opposite from the eccentricity axis  122 . The force vector in this 1-cylinder fuel pump  20  always acts exclusively in the direction of the cylinder axis, which as shown in  FIG. 11 , coincides with the eccentricity axis  122  at the top dead center. 
   The foregoing relates to preferred exemplary embodiments of the invention, it being understood that other variants and embodiments thereof are possible within the spirit and scope of the invention, the latter being defined by the appended claims.