EP1395753A1

EP1395753A1 - High-pressure pump for a fuel system of an internal combustion engine

Info

Publication number: EP1395753A1
Application number: EP02747167A
Authority: EP
Inventors: Helmut Rembold
Original assignee: Robert Bosch GmbH
Current assignee: Robert Bosch GmbH
Priority date: 2001-05-26
Filing date: 2002-05-24
Publication date: 2004-03-10
Anticipated expiration: 2022-05-24
Also published as: WO2002097268A1; EP1395753B1; US20040047746A1; DE50207940D1; US6889665B2

Abstract

The invention relates to a piston pump (20), in particular a high-pressure pump for a fuel system of an internal combustion engine, comprising a housing (72). Said pump is also provided with a piston (94), which delimits a working chamber (100). A drive shaft (78) is retained in the housing (72) by means of at least one shaft bearing and comprises at least one crank section (86). The piston (94) is at least indirectly supported on the crank section (86) of the drive shaft (78) using a piston bearing. A hydrostatic bearing system (62) is provided between parts of at least one of the bearings that can be displaced in relation to one another. Said bearing system is connected to the working chamber (100) via a fluidic connection. To improve the efficiency of the piston pump (20), a device (56; 118), which can temporarily interrupt the fluidic connection, is provided in said fluidic connection between the working chamber (100) and the hydrostatic bearing system (62).

Description

HIGH PRESSURE PUMP FOR A FUEL SYSTEM OF AN INTERNAL COMBUSTION ENGINE

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 which delimits 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 section , and with a piston bearing, via which the piston is supported at least indirectly on the crank section of the drive shaft, wherein at least one of the bearings has a hydrostatic bearing between parts which are movable relative to one another and is connected to the working space via a fluid connection.

Such a piston pump is known as a radial piston pump from DE 197 05 205 AI. With this radial piston pump, a race is placed on the eccentric section of a drive shaft. This has a flat contact surface on which a slide shoe of an axially reciprocating piston rests. Between the contact surface of the race and the slide shoe, there is a relief chamber which is connected to one of the axial holes in the slide shoe and in the piston Piston limited working space is connected. When the piston carries out a delivery stroke, the pressure in the working space increases, which is transmitted through the bore in the piston to the relief chamber and thus leads to a reduction in the contact force between the slide shoe and the race. A hydrostatic bearing is thus created by the relief chamber. This reduces friction and wear between the shoe and the race.

During operation of the known piston pump, however, it was found that the efficiency of this piston pump is good, but not yet optimal.

The present invention therefore has the task of developing a piston pump of the type mentioned at the outset in such a way that it has an even better efficiency.

In a piston pump of the type mentioned at the outset, this object is achieved in that a device is present in the fluid connection between the working space and the hydrostatic bearing, which device can temporarily interrupt the fluid connection.

Advantages of the invention

According to the invention, it was recognized that there is a leak in the area of the chamber between the parts that are movable relative to one another, ie fluid that is to be conveyed by the piston pump returns as leakage fluid via the hydrostatic bearing, for example to the inlet of the piston pump. This leakage affects the efficiency of the piston pump. Furthermore, it was recognized that relief of a bearing is not necessary at all times during a work cycle of the piston pump. Basically, a relief of the bearing parts lying against each other and movable relative to each other is only to those Times make sense when these two parts are pressed against each other with a relatively high force. In case of a

Piston pump is essentially the case during the delivery stroke.

Since, according to the invention, there is a device in the fluid connection between the working space and the hydrostatic bearing which can temporarily interrupt the fluid connection, the period of time for which fluid flows from the working space into the hydrostatic bearing can be limited to the necessary extent. As a result, the amount of leakage of fluid is reduced during the operation of the piston pump without the friction between parts of a bearing of the piston pump that are movable relative to one another being increased to an undesirable extent. This ultimately increases the efficiency of the piston pump without restricting the service life of the piston pump.

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

For example, it is proposed that the device which can temporarily interrupt the fluid connection comprises a pressure relief valve. This is installed in the fluid connection in such a way that it only releases it when the pressure in the area of the fluid connection facing the working area exceeds a limit value. This is based on the idea that the loads on the bearings are greatest when the pressure in the work area is high. Such a piston pump is simple to build and works reliably.

It is also possible that the device which can temporarily interrupt the fluid connection comprises a switching valve. In this development, any time can be selected at which the hydrostatic bearing is connected to the work space or at which the Connection is interrupted. This again allows a reduction in the amount of fluid used for the hydrostatic bearing.

It is particularly preferred if 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 towards the end of a delivery stroke, and the quantity of the effectively delivered fluid is thus limited. In this further development, hardly any fluid is lost for the hydrostatic bearing, since only the fluid is used for this which should not reach the actual outlet of the piston pump anyway in order to limit the delivery rate, but rather is returned to its inlet.

The piston pump according to the invention is relatively small when the device which can temporarily interrupt the fluid connection is accommodated in the piston. However, it is also possible to accommodate them in the piston pump housing. In this case, the device is more accessible for maintenance purposes, for example.

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

The hydrostatic bearing can comprise a chamber which is delimited in the azimuthal direction. This reduces the volume of the chamber and ultimately the amount of fluid required to form a hydrostatic bearing. Such a limitation of the chamber does not lead to a significant increase in the bearing frictional forces, since the hydrostatic bearing only has to act in the direction of the force peaks. These naturally occur predominantly when the piston is in the area of its top dead center, that is to say the fluid enclosed in the working space is 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 ° / 2xnumber of pistons.

The length and also the width of the chamber create the optimal 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. With such a pressure damper, the temporal course of the fluid flow, which flows from the work 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. If this quantity control valve is opened towards the end of the delivery stroke, there is an abrupt pressure increase in the fluid connection and thus also in the chamber. This pressure increase can be flattened somewhat by such a pressure damper.

The development in which at least one flow restrictor is present between the fluid connection and the pressure damper is aimed in the same direction. By means of such a flow restrictor, the pressure gradient over time in the fluid connection is, for example, when using a Pressure relief valve or a switching valve is reduced and the pressure rise stretched somewhat over time. The hydrostatic bearing is therefore available over a longer period of time than the fluid connection between the chamber and the work space is open.

The fluid connection to the chamber in the shaft bearing can include a flow channel in the housing, an annular groove connected to it in a bearing shell or in the shaft, a radial bore in the shaft connected to the annular groove, an axial bore in the shaft connected to this and an associated bore include radial bore in the shaft, which opens into the chamber in the shaft bearing. Such holes are easy to make, which facilitates the establishment of the fluid connection.

The same also applies to that fluid connection which leads to the chamber in the piston bearing and which comprises a radial bore which extends from the axial bore in the shaft and which opens 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 which delivers into a fuel manifold, and with at least one fuel injection device which is connected to the fuel manifold 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, it is proposed according to the invention that the fuel pump is designed in the above manner.

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

drawing

Exemplary embodiments of the invention are explained in detail below with reference to the accompanying drawing. The drawing shows:

1 shows a schematic basic illustration of a fuel system with a first exemplary embodiment of a fuel pump;

Fig. 2 is a partially sectioned illustration of the fuel pump of Fig. 1;

Fig. 3: a section along the line III-III of Fig. 2;

Fig. 4: a section along the line IV-IV of Fig. 2;

5: a representation of the angular range of a

Force vector of the fuel pump of Figure 2 based on the longitudinal axis of a drive shaft.

6: a representation similar to FIG. 1 of a fuel system with a second exemplary embodiment of a fuel pump;

FIG. 7: an illustration similar to FIG. 2 of the fuel pump from FIG. 6;

8: a representation similar to FIG. 1 of a fuel system with a third exemplary 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 FIG. 4 of the corresponding area of the fuel pump from FIG. 8; and

11: a representation of the angular range of a

Force vector of the fuel pump of Figure 8 related to the longitudinal axis of a drive shaft.

Description of the embodiments

In FIG. 1, a fuel system bears the overall reference number 10. It is part of an internal combustion engine 11 and includes 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 high-pressure fuel pump, designated overall by 20, which is driven by a crankshaft (not shown) of the internal combustion engine 11. Their exact structure is discussed in detail below.

A fuel line (without reference number) leads from an outlet 22 to a fuel collecting line 24, which is also commonly referred to as "aH". A plurality of fuel injection devices 26 are connected to the fuel collecting line 24. These are high pressure -Injection valves or injectors, the latter being one on the engine block (not shown)

Internal combustion engine (not shown) attached and inject the fuel directly into combustion chambers 28.

The pressure in the fuel collecting line 24 is detected by a pressure sensor 30, which delivers a corresponding signal to a control and regulating device 32. This, in turn, is to be represented on the output side in a manner to be described in more detail connected 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 shape. In principle, the high-pressure fuel pump 20 is constructed as follows:

A flow channel 34 leads from the inlet 18 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. Flow channels branch off from the branching point 38 to the individual cylinders 40a, 40b and 40c. The cylinders 40a-40c are constructed identically. For reasons of illustration, the reference symbols are only entered for one cylinder.

Each cylinder 40a-40c has a check valve 42 on the inlet side, 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 come together again 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 switching valve, which is open in its rest position 58 and 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, a flow channel 64 branches off from the flow channel 54 and ultimately between the Check valve 36 and the branch point 38 opens into the flow channel 34. A pressure damper 66, which in the present case is a spring / piston accumulator, is arranged in the flow channel 64. However, it is also possible to design the pressure damper 66 as a compression volume, bellows, membrane accumulator, etc. Upstream of the pressure damper 66 there is a first flow restrictor 68 in the flow channel 64, and a further flow restrictor 70 downstream of the pressure damper 66.

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

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

The drive shaft 78 is supported in the area of its two longitudinal ends by a bearing in the housing 72. The bearing on the left in FIG. 2 bears the reference symbol 80. In FIG. 2 to the right of the bearing 80, the horizontal recess 74 is sealed off from 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 section 86, on which a race 88 is placed.

The vertical recess 76 is through an upward Cover 90 closed. A guide sleeve 92 is inserted into the recess 76. In this in turn, a piston 94 is axially displaceably guided. A foot 96 is welded to the lower end of the piston 94 in FIG. 2. A compression spring 98 is tensioned between the foot 96 and the guide sleeve 92. This causes the foot 96 and thus ultimately the piston 94 to act against the race 88. The race 88 thus forms a piston bearing for the piston 94 relative to the drive shaft 78 (without reference numerals).

A working space 100 is formed above the piston 94 in FIG. 2. Coming from the left in FIG. 2 is the flow channel in which the check valve 42 is arranged. In FIG. 2 to the right of the working space 100 is the flow channel in which the check valve 46 is arranged. Neither the branching point 38 nor the collection point 48 are visible in the sectional plane shown in FIG. 2. The working space 100 and the piston 94 are part of the pump unit 44 of the cylinder 40 shown.

The hydrostatic bearing 62 is constructed as follows:

The flow channel 54 leads from the switching valve 56 to the horizontal recess 74. The flow channel 54 is continued via a bore 102 in the bearing shell 84 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 made in the drive shaft 78, which opens into an axial bore 108 in the drive shaft 78. This continues into the eccentric section 86 of the drive shaft 78.

From the axial bore 108, a radial bore 110 leads outwards to a recess (without reference number) on the outer lateral surface of the drive shaft 78. As can be seen from FIG. 3, this recess runs in an azimuthal direction Direction over an angular range of approximately 60 ° (only the shaft 78 and the bearing shell 84 are shown in FIG. 3 for reasons of illustration; in an exemplary embodiment not shown, the angle is less than 60 °). It forms a chamber 112, in which a hydrostatic counterforce to the forces originating from the piston 94 is generated in a manner still to be explained.

In the same manner, but offset by 180 °, a radial bore 114 branches off from the axial bore 108 in the region of the eccentric section 86 and opens out in a chamber 116 in an analogous manner. As can be seen from FIG. 4, this chamber 116 also extends in the azimuthal direction over an angular range of approximately 60 ° (in an exemplary embodiment not shown, this angle is less than 60 °). Here too, only the shaft 86 and the race 88 are shown in FIG. 4 for reasons of better illustration.

The high pressure fuel pump 20 operates as follows:

Due to the eccentric section 86, a rotation of the drive shaft 78 is converted into an axial back-and-forth movement of the piston 94. The switching valve 56 is controlled by the control and regulating device 32 in such a way that it is initially closed during a delivery stroke of the piston 94, when the piston 94 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 reaches the fuel collecting line 24 from the working space 100. When the desired pressure in the fuel collecting line 24 has been reached, this is detected by the pressure sensor 30.

The control and regulating device 32 then controls the switching valve 56 so that it opens. Thus, the fluid communication between the work space 100 and the chambers 112 and 116 becomes the hydrostatic bearing 62 opened. This increases the pressure in the chambers 112 and 116, which creates 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 the 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, as a result of which the fluid connection between the working space 100 and the two chambers 112 and 116 is interrupted again.

However, closing the switching valve 56 does not immediately end the hydrostatic counterforce that is generated in the chambers 112 and 116. On the one hand, it takes a certain time until, on the one hand, the fluid flows out 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 also delivers 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 any case in the case of a multi-cylinder pump a maximum of 360 ° / 2 x the number of cylinders, in this case 60 ° for three cylinders. This angular extension results from the following considerations:

As can be seen from FIG. 5, the force vector resulting from the pressure loading of the pistons of the cylinders 40a to 40c varies in a range of approximately 60 ° in the present three-cylinder high-pressure pump 20 from the angular position of the drive shaft 78. The beginning of the area is in turn offset by approximately 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, the force vector rotates synchronously with the drive shaft 78 about its longitudinal axis. Based on this load, the hydrostatic force on the piston bearing (race 88 and shaft 78) relieves the pressure in the area of the chamber 116 and on the shaft bearing (bearing shell 84 and shaft 78) by 180 ° in the area of the chamber 112.

In the illustrated in Figures 1 to 5

In the exemplary embodiment, the efficiency of the pump 10 is hardly impaired by the hydrostatic bearings 62, since fluid is used for its production, which fluid is used anyway for pressure control by the switching valve 56. There is therefore no additional leakage for the hydrostatic bearing.

6 and 7, a second embodiment of a high pressure fuel pump 20 is shown. Such parts, elements and areas which have equivalent functions to the parts, elements and areas already described above have the same reference numerals and are not explained again in detail.

In contrast to the exemplary embodiment described above, in the fluid connection 54 between the working space 100 and the chambers 112 and 116 there is not a switching valve but an overpressure valve 118. This only releases the fluid connection 54 when the pressure in the work space 100 exceeds a certain limit value. The hydrostatic counterforce therefore only becomes fully effective above the opening pressure of the pressure relief valve 118.

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

In addition to the first exemplary embodiment, an axially extending groove 120 is also present in the inside of the bearing shell 84. This leads from the space to the right of the bearing shell 84 to the space to the left of the bearing shell 84 in the recess 74. The groove 120 prevents the leakage between the drive shaft 78 and the bearing shell 84 from building up on the face side, which is impermissibly high Axial forces on the drive shaft 78 could cause. The space of the horizontal recess 74 to the left of the bearing shell 84 is connected to the inlet 18 of the high-pressure fuel pump 20 in a manner not shown here.

A further exemplary embodiment of a high-pressure fuel pump is shown in FIG. Here too, such components and areas whose function is equivalent to corresponding components and areas of the previous figures have the same reference numerals and are not explained again in detail.

In contrast to the exemplary embodiments shown in FIGS. 1 and 6, a 1-cylinder Piston pump 20 shown. Among other things, this leads to a different orientation of the chambers 112 and 116, as can be seen from FIGS. 9 and 10. Thereafter, the chamber 116 is formed in a range of approximately 60 ° on both sides of the eccentric axis 122. It therefore has approximately twice the angular extent than the corresponding chamber in the previous exemplary embodiments. Furthermore, it is offset from the previous exemplary 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 to say arranged with its central axis opposite to the eccentric axis 122. In this 1-cylinder fuel pump 20, the force vector always acts only in the direction of the cylinder axis, which, as shown in FIG. 11, coincides with the eccentric axis 122 at top dead center.

Claims

Expectations

1. Piston pump, in particular high-pressure pump (20) for a fuel system (10) of an internal combustion engine, with a housing (72), with at least one piston (94) that delimits a working space (100), with a drive shaft (78), which at least one shaft bearing is held in the housing (72) and has at least one crank section (86), and with a piston bearing, via which the piston (94) is supported at least indirectly on the crank section (86) of the shaft (78), wherein between there is a hydrostatic bearing (62) which is movable relative to one another in at least one of the bearings and is connected to the working space (100) via a fluid connection, characterized in that in the fluid connection between the working space (100) and the hydrostatic bearing (62) there is a device (56; 118) which can temporarily interrupt the fluid connection.

2. Piston pump (20) according to claim 1, characterized in that the device which can temporarily interrupt the fluid connection comprises a pressure relief valve (118).

3. Piston pump (20) according to one of claims 1 or 2, characterized in that the device which can temporarily interrupt the fluid connection comprises a switching valve (56).

4. Piston pump (20) according to claim 3, characterized in that that the switching valve is the quantity control valve (56) of the piston pump.

5. Piston pump (20) according to any one of the preceding claims, characterized in that the device which can temporarily interrupt the fluid connection is accommodated in the piston (94).

6. Piston pump (20) according to one of claims 1 to 4, characterized in that the device (56; 118), which can temporarily interrupt the fluid connection, in the housing

(72) is housed.

7. Piston pump (20) according to one of the preceding claims, characterized in that at least one hydrostatic bearing (62) is present in the piston bearing and in the shaft bearing.

8. Piston pump (20) according to one of the preceding claims, characterized in that the hydrostatic bearing (62) each comprises at least one chamber (112, 116) which is delimited in the azimuthal direction.

9. piston pump (20) according to claim 8, characterized in that it has a plurality of radially distributed pistons (94) that the angular range over which the chamber

(112, 116) extends in the azimuthal direction, is equal to or less than 360 ° / 2x the number of pistons (94), and that this region is offset by approximately 60 ° in the direction of rotation to a co-rotating axis (122) pointing in the eccentric direction.

10. Piston pump (20) according to one of the preceding claims, characterized in that the fluid connection is connected to a pressure damper (66).

11. Piston pump (20) according to claim 10, according to one of the preceding claims, characterized in that between the fluid connection and the pressure damper (66) at least one flow restrictor (68) is present.

12. Piston pump (20) according to one of claims 8 to 11, characterized in that the fluid connection to the chamber (112) in the shaft bearing has a flow channel (54) in the housing (72), an annular groove (104) connected thereto in one

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 this in the shaft (78), and one connected to this includes radial bore (110) in the shaft (78) which opens into the chamber (112) in the shaft bearing.

13. Piston pump (20) according to claim 12, characterized in that the fluid connection to the chamber (116) in the piston bearing comprises a radial bore (114) extending from the axial bore (108) in the shaft (78), which into the chamber ( 116) opens into the piston bearing.

14. Fuel system (10) for an internal combustion engine (11), with a fuel tank (12), a fuel pump (20), which promotes in a fuel manifold (24), and with at least one fuel injection device (26) is connected to the fuel manifold (24) and injects the fuel directly into the combustion chamber (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 13.

15. Internal combustion engine with at least one combustion chamber (28) into which the fuel is injected directly, characterized in that it has a fuel system (10) according to claim 14.