WO1998045594A1

WO1998045594A1 - Injection system, pressure valve, flow control valve, and method for setting the fuel pressure

Info

Publication number: WO1998045594A1
Application number: PCT/DE1998/000468
Authority: WO
Inventors: Hinrich KRÜGER; Martin Werner
Original assignee: Siemens Aktiengesellschaft
Priority date: 1997-04-08
Filing date: 1998-02-17
Publication date: 1998-10-15
Also published as: DE19714489C1; EP0974008A1; EP0974008B1

Abstract

Disclosed are a pressure valve and a flow control valve placed in a housing, which are operatively connected through a spring in such a way that the flow increase causes the clamping force from the pressure valve to also increase. Both the pressure valve and the flow control valve are controlled by means of an actuator.

Description

description

Injection system, pressure valve and volume flow control valve and method for regulating a fuel pressure

The invention relates to an injection system for an internal combustion engine according to the preamble of claim! ^', A volume flow regulating valve and pressure valve according to the preamble of claim 3 and a method for controlling egg nes fuel pressure in a fuel accumulator according to the preamble of claim 7.

The control of the fuel pressure in a fuel accumulator is particularly important in a common rail system, since the maximum fuel pressure in a common rail system is, for example, 1600 bar. Due to the high pressure, it is advantageous to regulate the pressure in the fuel accumulator with as little power loss as possible.

A system for regulating the fuel pressure in a fuel accumulator is already known, in which the fuel pump always pumps too much fuel into the fuel accumulator and a pressure valve opens when a predetermined fuel pressure is exceeded. However, this system has a relatively low efficiency.

Furthermore, it is known to provide a volume flow control valve on the inlet side of the high-pressure pump to improve the efficiency and to regulate the pressure in the fuel accumulator with the volume flow control valve. Here, however, it is necessary to additionally provide a pressure valve on the fuel accumulator, which can quickly lower the pressure in the fuel accumulator, which is necessary, for example, when changing from full load to idling mode. In addition, the pressure control valve required to depressurize the fuel accumulator after switching off the internal combustion engine.

A fuel injection system is known from US Pat. No. 4,884,545, in which a fuel pump conveys fuel into a fuel accumulator, which forwards the fuel to injection valves. A volume flow control valve is provided in the inlet to the fuel pump, which adjusts the fuel flow to the fuel pump. The volume flow control valve is controlled by a control unit via an actuator. A safety valve 13 is provided on the fuel accumulator, which allows fuel to flow back from the fuel accumulator to the fuel tank when a predetermined pressure is exceeded.

In the post-published patent application DE 196 12 413 AI, a fuel injection system is described in which a fuel pump supplies fuel to a fuel accumulator that supplies the fuel injectors to the fuel. A volume flow control valve is provided in the inlet to the fuel pump and is controlled by a control unit. The

The fuel accumulator is connected to a pressure control valve that is mechanically coupled to the volume flow control valve. The mechanical coupling is designed in such a way that the pressure control valve can be moved from a closed position into a through position by the movement of the actuator which controls the volume flow control valve, which leads to a rapid pressure relief of the fuel accumulator.

The object of the invention is an inexpensive

To provide pressure control for a fuel accumulator that is also highly efficient.

The object of the invention is achieved by the features of claims 1, 3 and 8. A major advantage of The invention is based on the fact that both the volume flow in the inlet to the high-pressure pump and the pressure in the fuel accumulator are regulated with a single control valve.

Further advantageous developments and improvements of the invention are specified in the dependent claims.

The invention is explained in more detail below with reference to the figures; show it:

1 shows an injection system with the control valve according to the invention, FIG. 2 shows the schematic structure of the control valve, FIG. 3 shows a further embodiment of the control valve, FIG. 4 shows a holding pressure and volume flow characteristic curve, and FIG. 5 shows a preferred design of the control valve.

FIG. 1 schematically shows the structure of an injection system that supplies fuel from a fuel tank 11 via a prefeed pump 2 via a control valve 10 to a high-pressure pump 1. The high-pressure pump 1 compresses the supplied fuel and releases the fuel under high pressure into the fuel accumulator 4. The fuel accumulator 4 is connected to injection valves 5, via which the fuel is injected into an internal combustion engine. A pre-pressure control valve 3 is connected in parallel with the pre-feed pump 2 and adjusts the fuel pressure after the pre-feed pump 2 to a predetermined value.

The fuel accumulator 4 is connected to the control valve 10 via a return line 26. The control valve 10 is also connected to a tank line 27 which is led to the fuel tank 11. A pressure sensor 9 is arranged on the fuel accumulator 4 and is connected to a control unit 6 via a signal line. In addition, a speed sensor 8 and a gas pedal sensor 7 is provided, which are also connected to the control unit 6 via a signal line. The control unit β has a data memory 28 and is connected to the injection valves 5 via first control lines and to the control valve 10 via a second control line.

The arrangement according to FIG. 1 functions as follows: The control device 6 controls the injection valves 5 according to a corresponding program, which is stored in the data memory 28, as a function of the speed of the internal combustion engine and the driver's request. In addition, the control unit 6 controls the control valve 10 as a function of the speed of the internal combustion engine and the fuel pressure in the fuel accumulator 4 and thus regulates the fuel pressure in the fuel accumulator 4.

Figure 2 shows schematically the structure of the control valve 10. The control valve 10 has an actuator 21 which is designed, for example, as a magnet. The actuator 21 is connected directly to a volume closing member 20, which connects the fuel inlet 24, which comes from the prefeed pump 2, with the fuel outlet 23, which is led to the high-pressure pump 1. In addition, the volume closing member 20 is connected via a spring 12 to a pressure closing member 22, which closes the connection between the return line 26 and the tank line 27 with an adjustable holding pressure.

In the rest position, the connection cross section between the fuel inlet 24 and the fuel outlet 23 is closed by the volume closing member 20 and the connection between the return line 26 and the tank line 27 is opened. If the control device 6 now controls the magnet 21, the volume closing member 20 is moved in the direction of the pressure closing member 22 and the connection cross section between the fuel inlet 24 and the fuel outlet 23 is thus opened. To- which is biased by the spring 12, the pressure closing member 22 against the opening cross section of the return line 26. The spring 12 is preferably designed such that, in the rest position, the pressure closing member 22 releases the return line 26 and the return line 26 is connected to the tank line 27. FIG. 3 shows a development of the control valve 10, in which the operative connection between the volume closing element 20 and the pressure closing element 22 is achieved via a first coupling spring 16 and a second coupling spring 17. The second coupling spring 17 is biased to a predetermined spring force. If the control device 6 now controls the actuator 21, the volume closing member 20 is moved and the volume flow that flows to the high-pressure pump is increased. In addition, the pressure closing member 22 is biased against the return line 26 via the first coupling spring 16. If the first coupling spring 16 is now compressed to such an extent that the pretensioned spring force of the second coupling spring 17 is reached, the first and second coupling springs 16, 17 act in series connection when the volume closing element 20 is deflected further.

FIG. 4 shows characteristic curves for the holding pressure P for various spring couplings between the volume closing element and the pressure closing element as a function of the displacement path S of the actuator 21 and as a function of the opening cross section Q that the volume closing element 20 opens. The holding pressure P corresponds to a closing force F. The deflection in the range greater than S1 corresponds to a volume flow Q> 0. However, the rest position of the actuator 21 is preferably s = 0 in the diagram, which ensures that the fuel accumulator is depressurized in the rest position, ie in the de-energized state of the actuator 21. When the actuator 21 is deflected between 0 and S1, the pressure closing member 22 first builds up a holding pressure before the volume closing member 20 establishes the connection cross section between the Fuel inlet 24 and fuel outlet 23 open at the first deflection S1.

The characteristic curve A corresponds to the control valve 10 of FIG. 2, in which only one spring 12 is provided between the volume closing element 20 and the pressure closing element 22. The pressure that is set by the pressure closing member 22 increases linearly with the deflection s of the actuator 21. It can be seen from Figure 4 that the volume closure member 20 initially, i.e. for the deflection s <Sl covers an empty path in which the fuel outlet is not yet connected to the fuel inlet. The pressure closing element 22 is biased with a holding force F0 at the first deflection S1, at which the volume closing element opens the connection cross section between the fuel feed line and the fuel discharge line. The linear characteristic has the disadvantage that a large holding force F is built up with a large deflection s.

The characteristic pressure curve for the control valve of FIG. 3, in which a first and a second coupling spring 16, 17 are provided between the volume closing element 20 and the pressure closing element 22, is shown in the characteristic curve B. In the rest position, the first coupling spring 16 is relaxed and the second coupling spring 17 is biased by means of a stop 18 and a transmission disk 19. Here, too, an empty travel is provided for the volume closing member 20, so that it only opens when the pressure closing member 22 is already biased against the return line 26 with a holding force F1. If the actuator 21 is now activated and the volume closing element 20 is deflected, the holding force with which the pressure closing element 22 is preloaded increases linearly up to a second deflection S2. The linear increase corresponds to the spring rate of the first coupling spring 16. From the second deflection S2, the first coupling spring 16 is tensioned such that the spring force of the first coupling spring 16 is the spring force of the pretensioned second coupling spring 17 reached. As soon as the tension of the first coupling spring 16 exceeds the pretensioning force of the second coupling spring 17, the transmission disk 19 is released from the stop 18 and the second coupling spring 17 is also compressed. The first and second coupling springs 16, 17 thus act in series from the second deflection S2. Therefore, from the second deflection S2, the linear increase in the holding force kinks into a second, flatter linear increase, which corresponds to a lower spring rate. From the second deflection S2, the holding force with which the

Pressure closing member 22 is biased per deflection unit s less than in the range between the deflection s = 0 and the second deflection S2, i.e. the spring rate is smaller for s> S2.

If, instead of the first and second coupling springs 16, 17, a spring, in particular a plate spring, is used, which has a degressive spring characteristic, the holding force of the pressure-closing element 22 is dependent on the deflection of the magnet 21 in accordance with the characteristic C of FIG. 4 the degressive spring characteristic increases the holding force of the pressure closing element 22 steeply with small deflections after the rest position s = 0 and changes into an almost horizontal course in the region of the second deflection S2. At the first deflection S1, the plate spring has a predetermined holding force F2.

The degressive disc spring or the two coupling springs 16, 17, starting from the rest position at s = 0, result in a steep increase in the holding force on the pressure closing element 22, which changes to a flat increase from a predeterminable, second deflection S2 of the actuator 21. The characteristic curves B and C are adapted to the actual ratios of the fuel pressure in the fuel accumulator and the volume flow supplied to the fuel accumulator. For a common Rail systems are already required with low volume flows, ie with a small amount of fuel that is injected and a low engine speed, high fuel pressures. Characteristic curves B and C thus offer good efficiency for electrical control, since unnecessarily high holding forces are avoided with large deflections. It is advantageous to install the spring with the degressive spring characteristic so that it is arranged at a predetermined distance from the volume closing element or pressure closing element, because this prevents a build-up of back pressure if the pressure in the fuel accumulator is reduced via the pressure closing element.

The characteristic curve forms B, C also offer the advantage that in the area between the rest position of the actuator 21 and the second deflection S2 there is a large change in the holding force of the

Pressure-closing member 22 is achieved with a small deflection of the actuator 21, with a small change in the deflection of the volume closing member 20 and thus with a small change in the volume flow. In this way, low volume flows can be set precisely. This can be seen from FIG. 4, since the pressure holding force F is a function, preferably proportional, of the control current I with which the actuator 21 is controlled. The volume flow Q is also preferably proportional to the deflection S.

Figure 5 shows a preferred embodiment of the control valve 10 with a spring combination corresponding to Figure 3. The control valve 10 has a valve body 31 which is screwed into a stepped bore of a housing 34 with a key attack 32 and a central thread 33. Housing 34 is preferably the housing of a high pressure pump. An inlet bore 35, an outlet bore 36, a high-pressure inlet bore 38 and a high-pressure outlet bore 37 are made in the housing 34. At the inlet bore 35, the force The material inlet 24, the fuel outlet 23 to the outlet bore 36, and the return line to the high-pressure inlet bore 38

26 and to the high pressure drain hole 37 is the tank line

27 connected. The inlet bore 35, the outlet bore 36 and the high-pressure outlet bore 37 are preferably designed as radial connection bores and open into a corresponding first annular channel 39, second annular channel 40 and third annular channel 41. In the example shown, the first annular channel 39 and the second annular channel 40 result from axially offset diameter steps in the housing 34 and in the valve body 31. The third annular channel 41 is introduced as a circumferential groove in the valve body 31. The control valve 10 shown in FIG. 5 is generally cylindrical symmetry with the axis of symmetry 71.

Between the first ring channel 39 and the central thread 33 there is a first sealing ring 42, between the first ring channel 39 and the second ring channel 40 there is a second sealing ring 43 and between the second ring channel 40 and the third ring channel 41 there is a third sealing ring 44 in the valve housing 31 brought in. The first, the second and the third sealing ring 42, 43, 44 are designed as radially sealing O-rings.

A first connecting bore 55, a second connecting bore 57 and a third connecting bore 64 are made in the valve body 31, starting from the first, the second and the third annular channel 39, 40, 41. The first, the second and the third connecting bore 55, 57, 64 connect the first, the second and the third ring channel 39, 40, 41 to a central bore 70 which is symmetrical to the axis of symmetry 71 and in the longitudinal direction of the valve body 31 in the valve body 31 is introduced. A control slide 53, which is designed as a sleeve, is introduced in the central bore 70 parallel to the axis of symmetry 71. In addition, a locking pin 51 is provided within the control slide 53 is arranged symmetrically and in the longitudinal direction to the axis of symmetry 71. The locking pin 51 and the control slide 53 are fitted into the central bore 70 and are arranged to be displaceable in the longitudinal direction of the central bore 70. The regulating slide 53 has an annular space 54 which is annular and open to the valve body 31 and which, in the rest position of the regulating slide 53, is only connected to the first inlet bore 35. If the control slide 53 is moved into the working position by the electromagnet 72, the first inlet bore 35 is connected to the first outlet bore 36 via the annular space 54. In this way, the volume flow that is supplied to the high pressure pump 1 is controlled.

The high-pressure inlet bore 38 is made in the center of the axis of symmetry 71 and at the lower end of the control valve 10. The central bore 70 is closed in the lower region by a head piece 45, which at the same time projects into the high-pressure inlet bore 38 with a closure piece. The head piece 45 is arranged centrally to the axis of symmetry 71 and has a pressure relief bore 48 in the center. The pressure relief bore 48 widens in the direction of the central bore 70 into a conical valve seat, in which a ball 50 is arranged, which is held by a receptacle 52 on the pressure relief bore 48. The receptacle 52 forms the lower end of the locking pin 51. Between the head piece 45 and the control slide 53, an annular space 63 is formed all around the locking pin 51, to which the high-pressure drain hole 37 is connected via the third connecting hole 64. The cylindrical extension of the head piece 45, which projects into the high-pressure inlet 38, is of an annular shape

Support ring 47 and a sealing ring 46, which seal the high pressure inlet bore 38.

The ball 50 is subjected to a corresponding holding force F by a corresponding movement of the locking pin 51. strikes, so that the pressure relief bore 48 is only released when the pressure in the pressure relief bore 48 is greater than the holding force F. In this way, the connection between the high pressure inlet 38 and the high pressure outlet 37 is controlled. The locking pin 51 with the ball 50 represents a pressure closing member 22 corresponding to Figures 2 and 3.

The sealing ring 46 and the support ring 47 offer the advantage that no axial contact force is required to seal the high-pressure inlet bore 38. The axial contact pressure would have to be absorbed between the head piece 45 and the central thread 33 by the valve housing 31 if, for example, a metallic flat seat or conical seat or a cutting ring would be used as a seal.

In the upper area of the control valve 10, an electromagnet 72 with a magnetic winding 73 and an associated armature guide rod 58 is arranged, which is guided in the middle of the axis of symmetry 71 in a guide sleeve 80. The armature guide rod 58 projects into the central bore 70 and is operatively connected to the control slide 53 via a second coupling spring 17. The second coupling spring 17 is prestressed by a transmission bush 60, the transmission bush 60 resting on a stop surface 62 of the control slide 53 and the second coupling spring 17 prestressing against a stop sleeve 74, which is designed as an end piece of the armature guide rod 58.

The transmission bushing 60 has, adjacent to the stop 62, a retaining ring 75 which is formed perpendicular to the axis of symmetry 71 and which, with a terminating bushing 76, which represents the upper end piece of the control slide 53, forms a sleeve-like second spring chamber 77 which, in the direction of the electromagnet 72, from the stop sleeve 74 is limited. The stop sleeve 74 is firmly connected to the terminating bush 76. This connection absorbs the biasing force of the second coupling spring 17.

The transmission socket 60 is terminated on the electromagnet side with an end plate 87. The locking pin 51, which terminates with a transmission plate 78, which has a diameter corresponding to the diameter of the transmission bush 60, projects into the transmission bush 60. The end plate 78 is arranged in the center of the axis of symmetry 71 and is arranged to be movable perpendicular to the axis of symmetry 71 and into the transmission bushing 60. A first coupling spring 16 is introduced between the transmission plate 78 and the end plate 87.

In the rest position, the armature guide rod 58 rests with a stop plate 79 against a stop 66 of the magnet housing 81. The stop 66 is preferably correspondingly adjustable, for example, using shims. The magnet armature 82 is adjusted via a further plate 85 and the stop plate 79 on the armature guide rod 58.

The locking pin 51 is not biased against the ball 50 in the rest position. Thus, the pressure relief bore 48 is in the rest position, i.e. opened without driving the electromagnet 72. The control slide 53 is arranged in the rest position so that the inlet bore 35 is connected to the annular space 54. However, the drain hole 36 is not connected to the annular space 54.

If the electromagnet 72 is now controlled by an output stage, the actuator guide rod 58 moves in the direction of the locking pin 51 and in doing so transmits a higher pressure holding force to the ball 50 via the second coupling spring 17 and the first coupling spring 16. near with the deflection S of the actuator guide rod 58, as shown in FIG.

At the same time, the control slide 53 is pushed directly from the actuator rod 58 in the direction of the second connecting bore 57. As soon as the control edge 56 of the annular space 54 reaches the second connection bore 57, a volume flow flows from the inlet bore 35 to the outlet bore 36. The volume flow to the high-pressure pump 1 is thus controlled. The second connecting bore 57 preferably has the shape of a rectangular longitudinal slot, which extends in the direction of movement of the control slide 53, so that the opening cross section is directly proportional to the path S with which the control slide 53 is moved. Another advantageous shape is a triangular opening cross section. The control slide 53 represents a volume closing member 20 according to FIGS. 2 and 3.

If the actuator guide rod 58 is now moved up to the second deflection S2, the first coupling spring 16 is compressed to such an extent that the spring force of the first coupling spring 16 is equal to the spring force of the second, preloaded coupling spring 17. With a further deflection of the actuator guide rod 58, both the first coupling spring 16 and the second coupling spring 17 are thus compressed.

In the area between the deflection 0 and the second deflection S2, the transmission bush 60 moves together with the stop sleeve 74 and the control slide 53. From the second deflection S2, the transmission bush 60 moves in the equilibrium of forces of the second coupling spring 17 and the first coupling spring 16 with respect to the stop sleeve 74 of the armature guide rod 58. In this way, the increase in the holding force F on the ball 50 kinks into one from the second deflection S2 smaller increase per deflection unit s, as can be seen from FIG. 4.

The high-pressure inlet bore 38 is only connected to the high-pressure outlet bore 37 when a greater pressure acts on the ball 50 from the high-pressure inlet bore than on the ball 50 due to the holding force F of the closing pin 51.

With the exception of the annular space 54, all spaces within the control valve are connected to the annular space 63 with the aid of grooves and bores, not shown, in order to enable displacement of the fuel when the individual parts are displaced.

Instead of the two coupling springs 16, 17, a degressive spring can also be used between the armature guide rod 58 and the locking pin 51. In this way, a pressure holding characteristic curve according to the characteristic curve C of FIG. 4 is made possible.

Claims

claims

1. injection system for an internal combustion engine with injection valves (5) which are connected to a fuel accumulator (4) which is connected to a fuel pump (1) which supplies the fuel accumulator (4) with fuel,

- With a pressure valve (10,22) on the fuel accumulator

(4) is connected, - with a pressure sensor (9), which is provided on the fuel accumulator (4) and measures the pressure in the fuel accumulator (4) and feeds it to a control unit (6), characterized in that

- That a volume flow control valve (10, 20) is provided in the inlet to the fuel pump (1), which adjusts the fuel flow to the fuel pump (1),

- That the control unit (6) controls the volume flow control valve (10, 20) via an actuator (21), and that the same actuator (21) adjusts the holding force of the pressure valve (10, 22).

2. Injection system according to claim 1, characterized in that the actuator (21) acts on the volume closing member (20) of the volume flow control valve (10, 20), that the volume closing member (20) in operative connection with the pressure closing member (22) of the pressure valve (10, 22) stands and the position of the volume closing element (20) defines the holding force of the pressure closing element (22) with which the pressure valve (10, 22) keeps the fuel accumulator closed.

3. Volume flow control valve and pressure valve with a valve body (31), into which a fuel inlet (35, 55) and a fuel outlet (36, 57) are introduced, with a control slide (53) as a volume closing element, which connects the cross-section between the fuel (35,55) and the fuel outlet (3657) controls, with a pressure inlet (48) and a pressure outlet

(64,41,37), which are introduced into the housing (31,45), with a pressure closing element (50, 52, 51), which closes the pressure inlet (48) with a specifiable holding force from the pressure outlet (64, 41, 37), with a controllable actuator (72, 58) that detects the position of the volume closing element (53) and the holding force of the pressure closing member (50,51,52) defines.

4. Volume flow and pressure valve according to claim 3, characterized in that the actuator (72,58) directly defines the volume connection member (53) in the position that the actuator (72,58) via spring means (16,17) Holding force of the pressure closing member (50,51,52) specifies.

5. Volume flow and pressure control valve according to claim 3, characterized in that the volume closing member (53)

Spring means (16, 17) the holding force of the pressure closing member (50, 51, 52) determines that the spring means (16, 17) have a spring rate which decreases with increasing deflection of the volume closing member (53).

6. Volume flow and pressure control valve according to claim 4 or 5, characterized in that a first and a second spring (16, 17) are arranged in series as spring means, so that starting from a rest position in which the first spring (16 ) relaxed and the second spring (17) is biased, only the first spring (16) up to a predetermined deflection (S2) and from the predetermined deflection (S2) the first and second springs (16, 17) are compressed in series .

7. Method for regulating the fuel pressure in a fuel accumulator (4) of an injection system of an internal combustion engine, in which the fuel accumulator (4) is supplied with fuel, in which the fuel flow to the fuel accumulator (4) is dependent, in particular, on the operating state of the internal combustion engine of the amount of fuel required for injection and of that for injection zen required fuel pressure is regulated, characterized in that a pressure valve (10,22) is provided which limits the pressure in the fuel accumulator (4) to a predetermined holding pressure, that the holding pressure of the pressure valve (10,22) is set as a function of the volume flow .

8. The method according to claim 7, characterized in that starting from a rest position, in which there is no pressure in the fuel accumulator (4) and no volumetric flow of the fuel pump (1) is supplied, first the holding pressure of the pressure valve (10, 22) to a specifiable , positive value is set, and that only then a volume flow of the fuel pump is supplied.

9. The method according to claim 7 or 8, characterized in that the holding pressure increases with increasing volume flow, preferably increases proportionally to the volume flow.

10. The method according to any one of claims 7 to 9, characterized in that the holding pressure increases starting from the rest position up to a predeterminable first volume flow with a first slope, that with a volume flow greater than the first volume flow, the holding pressure increases with increasing volume flow a second slope that is smaller than the first slope.

11. The method according to claim 7, characterized in that the holding pressure of the pressure valve is higher than the predetermined

Fuel pressure of the injection process is set.