EP1395753B1 - High-pressure pump for a fuel system of an internal combustion engine - Google Patents

Description

State of the art

The present invention initially relates to a piston pump, in particular a high pressure pump for a fuel system of an internal combustion engine, with a housing, with at least one piston defining a working space, with a drive shaft which is held in the housing via at least one shaft bearing and which has at least one crank portion , And with a piston bearing, via which the piston is supported at least indirectly on the crank portion of the drive shaft, wherein between relatively movable parts of at least one of the bearings, a hydrostatic bearing is present, which is connected via a fluid connection to the working space.

Such a piston pump is known as a radial piston pump from DE 197 05 205 A1. In this radial piston pump, a race is placed on the eccentric portion of a drive shaft. This has a flat contact surface on which a sliding shoe of an axially reciprocating piston rests. Between the contact surface of the race and the shoe a relief chamber is present, which via axial holes in the shoe and in the piston with a from the Piston limited work space is connected. When the piston performs a delivery stroke, the pressure in the working chamber increases, which is transmitted through the bore in the piston on the discharge chamber and in this way leads to a reduction in the contact force between the shoe and race. By the discharge chamber so a hydrostatic bearing is created. As a result, the friction and wear between the shoe and race is reduced.

It is also known from US 5 944 493 a radial piston pump, wherein the cylinder is connected via a check valve and a flow regulator with the piston bearings. In DE 199 00 564 A1, a common rail system is shown, in which a lubricant flow is branched off from a fuel flow delivered by a prefeed pump to a high-pressure pump. the pressure of the lubricant flow is adjusted by a pressure control valve.

The present invention therefore has the object, a piston pump of the type mentioned in such a way that it has a better efficiency.

This object is achieved in a piston pump of the type mentioned above in that in the fluid connection between the working space and the hydrostatic bearing, a switching valve is present, which can interrupt the fluid connection temporarily.

Advantages of the invention

According to the invention, it has been recognized that it is in the region of the chamber Leakage occurs between the relatively movable parts, ie fluid which is to be delivered by the piston pump passes as leakage fluid via the hydrostatic bearing, for example back to the inlet of the piston pump. This leakage affects the efficiency of the piston pump. Furthermore, it was recognized that a discharge of a bearing is not required at all times during a power stroke of the piston pump. Basically, a relief of the abutting and relatively movable bearing parts only makes sense at those times at which these two parts are pressed against each other with relatively high force. In the case of a piston pump, this is essentially the case during the delivery stroke.

According to the invention, there is a device in the fluid connection between the working space and the hydrostatic bearing which can intermittently interrupt the fluid connection, the period during which fluid flows from the working space into the hydrostatic bearing can be limited to the necessary extent. As a result, the leakage amount of fluid during operation of the piston pump is reduced, without the friction between relatively movable parts of a bearing of the piston pump would be increased to an undesirable extent. Thus, ultimately, the efficiency of the piston pump is increased, without the life of the piston pump is limited.

Thanks to the switching valve, any time points can be selected at which the hydrostatic bearing is connected to the working space or to which the connection is interrupted. This allows a reduction in the amount of fluid used for hydrostatic storage.

Particularly advantageous embodiments of the piston pump according to the invention are specified in subclaims.

For example, it is proposed that the switching valve is the quantity control valve of the piston pump. With such a quantity control valve, the outlet of the piston pump is usually short-circuited with its inlet toward the end of a delivery stroke, thus limiting the amount of fluid actually delivered. In this development, hardly any fluid is lost for the hydrostatic bearing, because for that only that fluid is used, which is not to reach the actual outlet of the piston pump anyway to limit the flow rate, but is returned to its inlet.

Relatively small, the piston pump according to the invention then builds when the device, which can interrupt the fluid connection temporarily, is housed in the piston. But it is also possible to accommodate them in the housing of the piston pump. In this case, the device is more accessible for maintenance purposes, for example.

Due to the significantly reduced amount of fluid which is required in the piston pump according to the invention for the production of a hydrostatic bearing, several, possibly all highly loaded bearings can be formed in the piston pump with such a hydrostatic bearing. This is taken into account by those training in which in each case at least one hydrostatic bearing is present in the piston bearing and the shaft bearing.

The hydrostatic bearing may comprise a chamber which is limited in the azimuthal direction. As a result, the volume of the chamber and thus ultimately the required amount of fluid for the formation of a hydrostatic bearing is reduced.

Such a limitation of the chamber does not lead to a significant increase in the bearing frictional forces, since the hydrostatic bearing must act only in the direction of the force peaks. These occur naturally predominantly when the piston is in the region of its top dead center, that is, the fluid enclosed in the working space is thus maximally compressed.

The piston pump according to the invention can be designed as a single and multi-cylinder piston pump. The angular range over which the chamber extends in the azimuthal direction is preferably less than 360 ° / 2x number of pistons.

The length and also the width of the chamber creates the optimum hydrostatic bearing for the individual application.

Another development is characterized in that the fluid connection is connected to a pressure damper. This can be used as compression volume, bellows, diaphragm accumulator or similar. be executed. By such a pressure damper, the time profile of the fluid flow, which flows from the working space to the chamber, can be designed. This is particularly advantageous if the device which can temporarily interrupt the fluid connection is the quantity control valve of the piston pump. When this quantity control valve is opened towards the end of the delivery stroke, there is a sudden pressure increase in the fluid connection and thus also in the chamber. By such a pressure damper, this pressure increase can be somewhat flattened.

In the same direction aims that development in which between the fluid connection and the pressure damper at least one flow restrictor is present. By such a flow restrictor, the temporal pressure gradient in the fluid connection, for example when using a Reduced pressure relief valve or a switching valve and the pressure rise slightly in time stretched. The hydrostatic bearing is thus available over a longer period of time than the fluid connection between the chamber and the working chamber is open.

The fluid connection to the chamber in the shaft bearing may include a flow channel in the housing, an annular groove in a bearing shell or shaft connected thereto, a radial bore in the shaft connected to the annular groove, an axial bore in the shaft connected thereto and one connected thereto include radial bore in the shaft, which opens into the chamber in the shaft bearing. Such holes are easy to incorporate, which facilitates the production of the fluid connection.

The same applies to those fluid connection, which leads to the chamber in the piston bearing and which comprises a projecting from the axial bore in the shaft radial bore, which opens into the chamber in the piston bearing.

The invention also relates to a fuel system for an internal combustion engine, comprising a fuel tank, a fuel pump which delivers into a fuel rail, and at least one fuel injector connected to the fuel rail and the fuel directly into the combustion chamber of an internal combustion engine injects.

In order to increase the efficiency of such a fuel system, the invention proposes that the fuel pump is formed in the above manner.

Furthermore, part of the invention is also an internal combustion engine with at least one combustion chamber, in which the fuel is injected directly. Such an internal combustion engine is advantageously provided with a fuel system of the above type.

drawing

Fig. 1:

eine schematische Prinzipdarstellung eines Kraftstoffsystems mit einem ersten Ausführungsbeispiel einer Kraftstoffpumpe;

Fig. 2:

eine teilweise geschnittene Darstellung der Kraftstoffpumpe von Fig. 1;

Fig. 3:

einen Schnitt längs der Linie III-III von Fig. 2;

Fig. 4:

einen Schnitt längs der Linie IV-IV von Fig. 2;

Fig. 5:

eine Darstellung des Winkelbereichs eines Kraftvektors der Kraftstoffpumpe von Fig. 2 bezogen auf die Längsachse einer Antriebswelle;

Fig. 6:

eine Darstellung ähnlich Fig. 1 eines Kraftstoffsystems mit einem zweiten Ausführungsbeispiel einer Kraftstoffpumpe;

Fig. 7:

eine Darstellung ähnlich Fig. 2 der Kraftstoffpumpe von Fig. 6;

Fig. 8:

eine Darstellung ähnlich Fig. 1 eines Kraftstoffsystems mit einem dritten Ausführungsbeispiel einer Kraftstoffpumpe;

Fig. 9:

eine Darstellung analog zu Fig. 3 des entsprechenden Bereichs der Kraftstoffpumpe von Fig. 8;

Fig. 10:

eine Darstellung analog zu Fig. 4 des entsprechenden Bereichs der Kraftstoffpumpe von Fig. 8; und

Fig. 11:

eine Darstellung des Winkelbereichs eines Kraftvektors der Kraftstoffpumpe von Fig. 8 bezogen auf die Längsachse einer Antriebswelle;

Hereinafter, embodiments of the invention will be explained in detail with reference to the accompanying drawings. In the drawing show:

Fig. 1:

a schematic diagram of a fuel system with a first embodiment of a fuel pump;

Fig. 2:

a partially sectioned view of the fuel pump of Fig. 1;

3:

a section along the line III-III of Fig. 2;

4:

a section along the line IV-IV of Fig. 2;

Fig. 5:

a representation of the angular range of a force vector of the fuel pump of Figure 2 with respect to the longitudinal axis of a drive shaft.

Fig. 6:

a representation similar to Figure 1 of a fuel system with a second embodiment of a fuel pump.

Fig. 7:

a representation similar to Figure 2 of the fuel pump of Fig. 6 .;

Fig. 8:

a representation similar to Figure 1 of a fuel system with a third embodiment of a fuel pump.

Fig. 9:

a representation analogous to FIG. 3 of the corresponding Area of the fuel pump of Fig. 8;

Fig. 10:

a representation analogous to Figure 4 of the corresponding portion of the fuel pump of Fig. 8. and

Fig. 11:

a representation of the angular range of a force vector of the fuel pump of Figure 8 with respect to the longitudinal axis of a drive shaft.

Description of the embodiments

In FIG. 1, a fuel system as a whole carries the reference numeral 10. It is part of an internal combustion engine 11 comprises a fuel reservoir 12, from which an electric fuel pump 14 conveys the fuel into a fuel line 16. This leads to an inlet 18 of a generally designated 20 high-pressure fuel pump, which is driven by a crankshaft, not shown, of the internal combustion engine 11. Their detailed structure will be discussed in detail below.

From an outlet 22, a fuel line (not numbered) leads to a fuel rail 24, commonly referred to as a "rail". To the fuel rail 24 a plurality of fuel injectors 26 are connected. These are high-pressure injectors or injectors. The latter are attached to the engine block (not shown) of an internal combustion engine (not shown) and inject the fuel directly into combustion chambers 28 a.

The pressure in the fuel rail 24 is detected by a pressure sensor 30, which supplies a corresponding signal to a control and regulating device 32. This, in turn, is on the output side in a manner to be shown in more detail the high-pressure fuel pump 20 is connected. The high-pressure fuel pump 20 is a radial piston pump with three radially arranged cylinders. In principle, the high pressure fuel pump 20 is constructed as follows:

From the inlet 18, a flow channel 34 leads via a check valve 36 to a branch point 38. The check valve 36 opens inwards and thus keeps pressure surges away from the fuel line 16 and the electric fuel pump 14. From the branching point 38, flow channels branch off to the individual cylinders 40a, 40b and 40c. The cylinders 40a-40c are constructed identically. For reasons of illustration, the reference numerals are entered only for one cylinder.

Each cylinder 40a-40c has on the input side via a check valve 42, a pump unit 44 and a check valve 46 arranged downstream of the pump unit 44. Downstream of the check valves 46, the flow channels of the individual cylinders 40a-40c again come together at a collection point 48. From there, a flow channel 50 leads via a further check valve 52 to the outlet 22 of the high-pressure fuel pump 20.

Between the collection point 48 and the check valve 52, a flow channel 54 branches off from the flow channel 50, in which a switching valve 56 is arranged. This is an electrically operated two / two-way valve, which is open in its rest position 58 and is closed in its actuated position 60. The switching valve 56 is controlled by the control and regulating device 32. The flow channel 54 leads from the switching valve 56 to a hydrostatic bearing 62, which is explained in detail below.

Downstream of the switching valve 56 branches off from the flow channel 54 from a flow channel 64, which ultimately between the Check valve 36 and the branch point 38 opens into the flow channel 34. In the flow channel 64, a pressure damper 66 is arranged, which in the present case is a spring / piston accumulator. However, the design of the pressure damper 66 as a compression volume, bellows, diaphragm accumulator, etc. is also possible. Upstream of the pressure damper 66 there is a first flow restrictor 68 in the flow channel 64, and downstream of the pressure damper 66 a further flow restrictor 70.

The exact configuration of the high-pressure fuel pump 20 can be seen in FIGS. 2-4. It should be noted that in this sectional plane, only one cylinder 40 is shown and individual channels, etc. are not visible.

The high-pressure fuel pump 20 includes a housing 72. In this a blind hole-like recess 74 is present, the longitudinal axis in Fig. 2 is horizontal. Further, a further recess 76 is introduced into the housing 72, which extends vertically in Fig. 2 and from the upper edge of the housing 72 extends into the horizontal recess 74 into it. In the horizontal recess 74, a drive shaft 78 is received. This is connected to the crankshaft (not shown) of the internal combustion engine.

The drive shaft 78 is mounted in the region of its two longitudinal ends in each case by a bearing in the housing 72. The left-hand bearing in FIG. 2 bears the reference numeral 80. In FIG. 2, to the right of the bearing 80, the horizontal recess 74 is sealed to the outside by a shaft seal 82. The right end of the drive shaft 78 is mounted in a hollow cylindrical bearing shell 84, which forms a shaft bearing. Approximately in its axial center, the drive shaft 78 has an eccentric portion 86, on which a race 88 is placed.

The vertical recess 76 is at the top by a Lid 90 closed. In the recess 76, a guide sleeve 92 is inserted. In this turn, a piston 94 is guided axially displaceable. At the bottom in Fig. 2 end of the piston 94, a foot 96 is welded. Between the foot 96 and the guide sleeve 92, a compression spring 98 is tensioned. By this the foot 96 and thus ultimately the piston 94 is acted upon against the race 88. The race 88 thus forms for the piston 94 relative to the drive shaft 78, a piston bearing (without reference numerals).

In Fig. 2 above the piston 94, a working space 100 is formed. In these opens in Fig. 2 coming from the left that flow channel in which the check valve 42 is arranged. In Fig. 2 to the right of the working space 100 that flow channel runs in which the check valve 46 is arranged. Neither the branching point 38 nor the collecting point 48 are visible in the sectional plane shown in FIG. The working space 100 and the piston 94 are part of the pumping unit 44 of the illustrated cylinder 40.

The hydrostatic bearing 62 is constructed as follows:

From the switching valve 56, the flow channel 54 leads to the horizontal recess 74. Via a bore 102 in the bearing shell 84, the flow channel 54 is continued up to an annular groove 104 on the inside of the bearing shell 84. At the same axial height as the annular groove 104, a radial bore 106 is introduced into the drive shaft 78, which opens into an axial bore 108 in the drive shaft 78. This continues into the eccentric portion 86 of the drive shaft 78 continues.

From the axial bore 108, a radial bore 110 leads outwardly to a recess (without reference numeral) on the outer circumferential surface of the drive shaft 78. This recess extends, as can be seen in Fig. 3, in azimuthal Direction over an angular range of approximately 60 ° (in Figure 3, for purposes of illustration, only the shaft 78 and the bearing shell 84 are shown; in an embodiment, not shown, the angle is less than 60 °). Through them, a chamber 112 is formed, in the manner to be explained, a hydrostatic counterforce is generated to the forces resulting from the piston 94.

In the same manner, but offset by 180 °, branches from the axial bore 108 in the region of the eccentric portion 86 from a radial bore 114 to the outside, which opens in a similar manner in a chamber 116. Also, this chamber 116 extends, as shown in FIG. 4, in the azimuthal direction over an angular range of about 60 ° (in an embodiment not shown, this angle is less than 60 °). Again, only the shaft 86 and the race 88 is shown in Figure 4 for the sake of better illustration.

The high pressure fuel pump 20 operates as follows:

Due to the eccentric portion 86, rotation of the drive shaft 78 is translated into axial reciprocation of the piston 94. The switching valve 56 is controlled by the control and regulating device 32 so that it is initially closed during a delivery stroke of the piston 94 when it moves upward. As a result, the pressure of the fluid enclosed in the working space 100 increases considerably. Via the flow channel 50, which is not visible in FIG. 2, the compressed fluid passes from the working space 100 into the fuel collecting line 24. When the desired pressure in the fuel collecting line 24 is reached, this is detected by the pressure sensor 30.

The control and regulating device 32 then controls the switching valve 56 so that this opens. Thus, the fluid communication between the working space 100 and the chambers 112 and 116 of the hydrostatic bearing 62 open. This increases the pressure in the chambers 112 and 116, creating a hydrostatic counterforce between the bearing shell 84 and the drive shaft 78 (shaft bearing) and on the other hand between the race 88 and drive shaft 78 (piston bearing) in the desired direction. At the end of the delivery stroke, the switching valve 56 is closed again by the control and regulating device 32, whereby the fluid connection between the working chamber 100 and the two chambers 112 and 116 is interrupted again.

However, closing the switching valve 56 does not immediately stop the hydrostatic drag generated in the chambers 112 and 116. On the one hand, it takes a certain amount of time until the fluid has flowed away on the one hand through the gap between the drive shaft 78 and the bearing shell 84 and on the other hand between the drive shaft 78 and the race 88. On the other hand, the pressure damper 66 acts as a pressure accumulator, which still promotes a certain amount of fluid into the chambers 112 and 116 even when the switching valve 56 is closed.

The time course of the hydrostatic counterforce generated by the pressure build-up in the chambers 112 and 116 is adjusted on the one hand by the width and the azimuthal angular extent of the chambers 112 and 116 and on the other hand by the properties of the pressure damper 66 and the two flow restrictors 68 and 70. The azimuthal angular extent of the chambers 112 and 116 is, as already mentioned, a maximum of 60 °, in each case in a multi-cylinder pump a maximum of 360 ° / 2 x number of cylinders, with three cylinders here so 60 °. This angular extension results from the following considerations:

As is apparent from Fig. 5, the force vector resulting from the pressure load of the pistons of the cylinders 40a to 40c in the present three-cylinder high-pressure pump 20 varies in a range of approximately 60 °, depending The beginning of the area is again offset by about 60 ° in the direction of rotation (arrow 121 in FIGS. 4 and 5) to a co-rotating axis 122 pointing in the eccentric direction. Within the said angular range of the force vector rotates synchronously with the drive shaft 78 about its longitudinal axis. Based on this load, the discharge is effected by the hydrostatic force on the piston bearing (race 88 and shaft 78) in the region of the chamber 116 and the shaft bearing (bearing shell 84 and shaft 78) offset by 180 ° in the region of the chamber 112th

In the embodiment shown in Figures 1 to 5, the efficiency of the pump 10 is hardly affected by the hydrostatic bearing 62, since fluid is used for their production, which is anyway used for pressure control by the switching valve 56. An additional leakage for the hydrostatic bearing is therefore not available.

FIGS. 6 and 7 show a second exemplary embodiment of a high-pressure fuel pump 20. Such parts, elements and regions, which have equivalent functions to parts, elements and regions already described above bear the same reference numerals and are not explained again in detail.

In contrast to the embodiment described above, in the fluid connection 54 between the working chamber 100 and the chambers 112 and 116, no switching valve, but a pressure relief valve 118 is arranged. This releases the fluid connection 54 only when the pressure in the working space 100 exceeds a certain limit. The hydrostatic counterforce is therefore only above the opening pressure of the pressure relief valve 118 fully effective.

The advantage is that - without the need for a electrical control - at low pressures in the working chamber 100 no fluid through the chambers 112 and 116 and the corresponding bearing gaps between the drive shaft 78 and bearing shell 84 on the one hand and between the drive shaft 78 and race 88 on the other hand occurs as leakage, resulting in a higher volumetric efficiency of the high pressure fuel pump 20th entails. Although a higher leakage occurs in the upper pressure range, it is at least compensated based on the overall efficiency due to the lower bearing load and the resulting higher mechanical efficiency. Regardless of the efficiency results in any case a significantly better service life of the high-pressure fuel pump 20th

In addition to the first embodiment, an axially extending groove 120 is still present in the inside of the bearing shell 84. This leads from the space present on the right of the bearing shell 84 to the space left in the recess 74 from the bearing shell 84. The groove 120 prevents pressure from building up over the leakage between the drive shaft 78 and the bearing shell 84 at the front end, which is unacceptably high Could cause axial forces on the drive shaft 78. The left of the bearing shell 84 existing space of the horizontal recess 74 is connected in a manner not shown here with the inlet 18 of the high-pressure fuel pump 20.

FIG. 8 shows a further exemplary embodiment of a high-pressure fuel pump. Again, bear such components and areas whose function is equivalent to corresponding components and areas of the preceding figures, the same reference numerals and are not explained again in detail.

In contrast to the exemplary embodiments illustrated in FIGS. 1 and 6, FIG. 8 shows a 1-cylinder piston pump 20 shown. This leads inter alia to a different orientation of the chambers 112 and 116, as shown in Figures 9 and 10 can be seen. Thereafter, the chamber 116 is formed in a range of about 60 ° on both sides of the eccentric axis 122. So it has approximately twice the angular extent as the corresponding chamber in the previous embodiments. Further, it is arranged offset from the previous embodiments by 90 ° counter to the direction of rotation of the drive shaft 78. The chamber 112 is offset from the chamber 116 by 180 °, that is arranged with its central axis opposite to the eccentric axis 122. The force vector acts in this 1-cylinder fuel pump 20 only in the direction of the cylinder axis, which coincides with the eccentric axis 122 as shown in FIG. 11 at top dead center.

Claims (12)

1. Piston pump, in particular high-pressure pump (20) for a fuel system (10) of an internal combustion engine, with a casing (72), with at least one piston (94) which delimits a working space (100), with a drive shaft (78) which is held in the housing (72) via at least one shaft bearing and which has at least one crank portion (86), and with a piston bearing, via which the piston (94) is supported at least indirectly on the crank portion (86) of the shaft (78), a hydrostatic mounting (62), which is connected to the working space (100) via a fluid connection, being present between those parts of at least one of the bearings which are movable in relation to one another, a device (56; 118) which can temporarily interrupt the fluid connection being present in the fluid connection between the working space (100) and the hydrostatic mounting (62), characterized in that the device which can temporarily interrupt the fluid connection comprises a switching valve (56).
2. Piston pump (20) according to Claim 1, characterized in that the switching valve is the quantity control valve (56) of the piston pump.
3. Piston pump (20) according to either one of Claims 1 and 2, characterized in that the device which can temporarily interrupt the fluid connection is accommodated in the piston (94).
4. Piston pump (20) according to one of the preceding claims, characterized in that the device (56; 118) which can temporarily interrupt the fluid connection is accommodated in the casing (72).
5. Piston pump (20) according to one of the preceding claims, characterized in that at least one hydrostatic mounting (62) is present in each case in the piston bearing and in the shaft bearing.
6. Piston pump (20) according to one of the preceding claims, characterized in that the hydrostatic mounting (62) comprises in each case at least one chamber (112, 116) which is delimited in the azimuthal direction.
7. Piston pump (20) according to Claim 8, characterized in that it has a plurality of radially distributed pistons (94), in that the angular range over which the chamber (112, 116) extends in the azimuthal direction is equal to or smaller than 360°/2 x the number of pistons (94), and in that this range is offset at approximately 60° in the direction of rotation with respect to a co-rotating axis (122) pointing in the eccentric direction.
8. Piston pump (20) according to one of the preceding claims, characterized in that the fluid connection is connected to a pressure damper (66).
9. Piston pump (20) according to Claim 8, characterized in that at least one flow throttle (68) is present between the fluid connection and the pressure damper (66).
10. Piston pump (20) according to one of Claims 6 to 9, characterized in that the fluid connection to the chamber (112) in the shaft bearing comprises a flow duct (54) in the casing (72), an annular groove (104), connected to the said flow duct, in a bearing shell (84) or in the shaft, a radial bore (106), connected to the annular groove (104), in the shaft (78), an axial bore (108), connected to the said radial bore, in the shaft (78), and a radial bore (110), connected to the said axial bore, in the shaft (78), which radial bore issues into the chamber (112) of the shaft bearing.
11. Piston pump (20) according to Claim 10, characterized in that the fluid connection to the chamber (116) in the piston bearing comprises a radial bore (114) which emanates from the axial bore (108) in the shaft (78) and which issues into the chamber (116) in the piston bearing.
12. Fuel system (10) for an internal combustion engine (11), with a fuel tank (12), with a fuel pump (20) which conveys into a fuel collecting line (24), and with at least one fuel injection device (26) which is connected to the fuel collecting line (24) and injects the fuel directly into the combustion space (28) of the internal combustion engine (11), characterized in that the high-pressure fuel pump (20) is designed according to one of Claims 1 to 11.

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