EP1654455A1

EP1654455A1 - Control valve for a fuel injector that contains a pressure intensifier

Info

Publication number: EP1654455A1
Application number: EP04738749A
Authority: EP
Inventors: Hans-Christoph Magel
Original assignee: Robert Bosch GmbH
Current assignee: Robert Bosch GmbH
Priority date: 2003-08-01
Filing date: 2004-06-22
Publication date: 2006-05-10
Anticipated expiration: 2024-06-22
Also published as: DE10335340A1; US7278398B2; CN100414090C; US20060196474A1; CN1833102A; WO2005015001A1; EP1654455B1; DE502004003161D1

Abstract

The invention relates to a servo valve for actuating a pressure intensifier (2), to which a fuel injector (9) is assigned. Said pressure intensifier (2) comprises a working space (4), which is separated from a differential pressure space (5) by an intensifier plunger (3, 50), and the pressure change in the differential pressure space (5) of the pressure intensifier (2) ensues via the servo valve (23), to which a switching valve (24) that activates this servo valve is assigned. Inside a valve housing (47; 61, 62, 63) of the servo valve (23), the control space (29) thereof can both be connected to a high-pressure source (1) as well as can be relieved from pressure in a low pressure-side return (25), and a pressure stage (44, 28), which acts in the closing direction of the valve plunger (32, 33, 60), is formed for effecting a rapid closing motion on the valve piston (32; 33, 60) between the control space (29) and a hydraulic space (42). In addition, control edges (36, 40) without a common opening phase are formed on the valve housing (47, 61, 62, 63) on the area of the valve plunger (32, 33, 60).

Description

Control valve for a fuel injector containing a pressure intensifier

Technical field

Both pressure-controlled and stroke-controlled injection systems can be used to supply the combustion chambers of auto-ignition internal combustion engines with fuel. In addition to pump-injector units, pump-line-injector units, accumulator injection systems are also used as fuel injection systems. Accumulator injection systems (Cornmon Rail) advantageously make it possible to adapt the injection pressure to the load and speed of the internal combustion engine. In order to achieve high specific outputs and to reduce the emissions of the internal combustion engine, the highest possible injection pressure is generally required.

State of the art

For reasons of strength, the achievable level of drain in storage injection systems used today is currently limited to approximately 1600 bar. This pressure intensifier is used to further increase the pressure on accumulator injection systems.

DE 101 23 910.6 relates to a fuel injection device with which fuel is supplied to the combustion chambers of a multi-cylinder internal combustion engine. The combustion engine's combustion chambers are each supplied with fuel via fuel injectors. These are acted upon by a high pressure source; Furthermore, the fuel injection device according to DE 101 23 910.6 comprises a pressure intensifier which has a movable pressure intensifier piston which separates a space which can be connected to the high pressure source from a high pressure space connected to the fuel injector. The fuel pressure in the high-pressure chamber can be varied by infecting a rear chamber of the pressure booster with fuel or by emptying this rear chamber of fuel. With the activation of the pressure booster via its rear space, it can be achieved that the activation losses in the high-pressure fuel system can be kept smaller in comparison to activation via a work space that is temporarily connected to the high-pressure source. Furthermore, the high pressure chamber of the pressure booster can only be relieved up to the drain level of the high pressure storage chamber and not up to the leakage pressure level. On the one hand, the hydraulic efficiency can be improved, on the other hand, a faster pressure build-up can be achieved down to the system pressure level, so that the time intervals between the individual injection phases can be shortened considerably.

A pressure intensifier can be used on each K-fuel injector of a combustion engine to increase the injection pressure. If the pressure intensifier is not activated, there is a flow connection from the drain accumulator to the injection nozzle. Such a system can be equipped with two valves with actuators that can be activated independently in order to ensure flexible injection course shaping. A disadvantage of this solution is the relatively high production outlay for controlling such a fuel injection system with two valves and two actuators which can be activated independently. The formation of a drain intensifier control valve can make it necessary to use a servo-hydraulically assisted valve due to the high discharge quantities from the differential pressure chamber of the pressure intensifier. However, this goes hand in hand with relatively high manufacturing costs. If, on the other hand, slide valves are used in such systems, this offers the advantage of lower manufacturing costs and a reduced tolerance tolerance. However, to ensure sufficient high-pressure tightness, a high overlap of the slide control edges must be ensured, which in turn requires a large valve lift of a few millimeters of the slide valve. This in turn means that an exact and rapid closing movement of a valve piston is difficult to achieve with such a solution, since the large spring forces required to bring about an exact and rapid closing movement cannot be achieved in the injector installation space. The large stroke of a valve piston designed as a slide valve requires a large installation space in order to achieve high spring forces.

Presentation of the invention

In order to ensure an exact fast closing movement of a control valve for a pressure intensifier, the control valve is designed as a slide valve with a pressure stage. The valve piston of the slide valve proposed according to the invention can be constructed in two parts, so that it has no double guide and can be manufactured relatively easily. Only two guides of different diameters are required. The separation point of the two-part valve piston is in a low-pressure chamber, whereas both ends of the valve piston parts are each subjected to high pressure, so that separation of the valve pistons is impossible. Due to the drainage stage formed on the slide valve, the valve is closed by hydraulic forces, so that the generation of a large spring force is not necessary. This in turn offers the advantage that the valve proposed according to the invention has no problems can be accommodated in the tree space available for fuel injectors.

A hydraulic restoring force can advantageously be generated via the pressure stage. In known slide valves with pressure stages there are several leakage paths, with a large pressure difference between rail pressure (system pressure) and low pressure being present on several guide sections of a servo valve piston. As a result, large overlap lengths must be provided for the guide sections in order to keep the amount of leakage within limits, which, according to this solution, results in long overall lengths of the servo valve piston.

If a servo valve piston is formed with only one guide section, which is pressurized with system pressure (rail pressure) when the fuel injector is at rest, the leakage can be reduced considerably. This one guide section has a smaller sealing diameter, since no valve pockets for the connection for control bores are to be provided in this section. The shorter overall length of the guide section of the servo piston also makes production easier.

As an alternative to designing the control valve as a 3/2-spool-spool valve with only one guide section that is pressurized with rail pressure when the fuel injector is at rest, an additional valve seat can be used to further reduce leakage losses. This additional valve seat can be designed as a flat seat and is structurally simple within a two-part valve housing and can be represented inexpensively in terms of production costs. In addition, the efficiency of a fuel injector can be significantly increased when using a 3/2-way slide valve with a flat seat as a control valve for the pressure intensifier. The required guide lengths and the valve stroke can be further reduced, which overall contributes to the reduction in installation space of the proposed 3/2 spool valve. This ensures the use of the solution according to the invention in the target installation space of modern internal combustion engines, where only a small installation space is available. The design of the servo valve as a 3/2-spool-spool valve with a flat seat enables the realization of a leak-free servo piston, with which a predefinable switching sequence for the valve closing can still be realized in order to enable secondary injection at an increased pressure level.

It applies to all variants of the servo valve proposed according to the invention that two control edges are used to control the pressure intensifier. The control edges (slide seal) are designed so that when closing there is a lateral delay between closing one and opening the other of the control edges, which is used for the construction of a drain cushion. Drawing Based on the drawing, the invention is explained in more detail below.

It shows:

FIG. 1 shows a first embodiment variant of a servo valve with a draining stage for controlling a pressure intensifier of a fuel injector,

FIG. 2 shows an embodiment variant of the servo valve shown in FIG. 1, designed as a slide valve, with a further hydraulic space acted upon via the differential pressure,

Figure 3 shows a further embodiment of a servo valve designed as a slide valve for controlling a pressure booster - shown in the idle state

FIG. 4 shows the embodiment variant shown in FIG. 3 of a servo valve designed as a slide valve with the pressure booster activated,

FIG. 5 shows a further embodiment of a servo valve designed as a slide valve with a multi-part servo valve housing and a flat seat formed therein in the idle state and

6 shows the embodiment variant shown in FIG. 5 of a servo valve designed as a slide valve with the pressure booster activated.

variants

FIG. 1 shows a servo valve designed as a slide valve for controlling a pressure intensifier on a fuel injector.

Via a high-pressure source 1, which can be a high-pressure collecting space (compression rail) or a high-pressure fuel pump, a pressure booster 2 is acted upon by fuel under high pressure. The pressure intensifier 2 comprises a working space 4 and a differential pressure space 5 (rear space), which are others are separated by a booster piston 3. The pressure booster 2 also includes a compression chamber 6. A high-pressure line 8 branches off from this, a check valve 7 being accommodated in the refilling branch of the pressure booster 2.

Via the high-pressure line 8, a fuel injector 9 is pressurized with fuel under a translated pressure, corresponding to the gear ratio of the pressure booster 2. The high-pressure line 8 merges into a nozzle chamber inlet 15, via which a nozzle chamber 14 is supplied with fuel. A first inlet throttle 12 branches off from the high-pressure line 8 into a control chamber 11. The control chamber 11 can be relieved of pressure via a first outlet throttle 13 when a first switching valve 18 is actuated into a first return 19 on the low pressure side. The lifting movement of an injection valve member 10, for example, in the form of a needle, is controlled by the pressurization or pressure relief of the control chamber 11. The injection valve member 10 comprises a pressure stage 17 in the area of the nozzle chamber 14. In addition, the injection valve member 10 is acted upon by a spring element 20 in the closing direction. The spring element 20 is arranged in a space of the injector body of the fuel injector 9, from which a second return 21 branches off on the low pressure side. When the injection valve member 10 is opened, the injection openings 16 opening into a combustion chamber of an internal combustion engine, which is not shown here, are released, so that fuel under high pressure can be injected into the combustion chamber of the internal combustion engine.

A control chamber 29 of a servo valve 23 is also supplied with high-pressure fuel from the high-pressure source 1 via a supply line 22. The servo valve 23 can be actuated by actuating a switching valve 24, which opens on the outlet side into a third return line 25 on the low pressure side. A second discharge throttle 27 can be connected between the second switching valve 24 and the control chamber 29 of the servo valve 23. A stop 30 for an end face 28 of a second servo valve piston 33 is also received in the control chamber 29. In the exemplary embodiment of a servo valve shown in FIG. 1, a first piston 32 and a second piston 33 are accommodated in the housing of the servo valve 23. The second piston 33 has a larger diameter compared to the diameter of the first piston 32. The second piston 33 is acted upon by a valve spring 31 accommodated in the control chamber 29 of the servo valve 23.

Below the second piston 33 there is a first hydraulic chamber 34 in the valve housing of the servo valve 23, which has a branch to a fourth return 35 on the low-pressure side. There is a below the first hydraulic space 34 second hydraulic space 38, which is hydraulically connected to the differential pressure space 5 (rear space) of the pressure booster 2 via a connecting line 43. Between the second hydraulic space 38 and a third hydraulic space 42, the first piston 32 has an asymmetrically designed section. This section is formed with an overlap length 41 and releases a flow cross section from the second hydraulic space 38 into the third hydraulic space 42. In the upper area of the first piston 32, this has a first overlap length 37 (hj) below the contact surface on the lower end face of the second piston 33. In the area of the first hydraulic chamber 34, a pressure stage is formed by the difference in diameter of the second piston 33 and the first piston 32, which pressure stage lies above a first sealing seat 36. At the lower area of the first piston 32, a sealing edge 40 is formed as a slide seat on the valve housing side. The hydraulic space 42 is supplied with fuel under high pressure via an overflow line 39, which branches off from the supply line 22 for filling the control space 29 of the servo valve 23. The end face of the first piston 32 enclosed by the third hydraulic space 42 is identified by reference numeral 44.

FIG. 2 shows a modification of the fuel injection system shown in FIG. 1, comprising a pressure intensifier and a fuel injector.

In contrast to the illustration according to FIG. 1, a connecting line section 46 branches off from the connecting line 43 of the differential pressure chamber 5 (rear chamber) of the pressure intensifier 2 in order to act on the second hydraulic chamber 38. The connecting line section 46 acts on a fourth hydraulic chamber 45 with fuel, which is under the pressure prevailing in the differential pressure chamber 5 (rear chamber) of the pressure converter 2. In comparison to the design of the first piston 32 in the embodiment variant shown in FIG. 1, the first piston 32 is formed with an extended length which penetrates the third hydraulic space 42. The end face 44 of the first piston 32 projects into the fourth hydraulic chamber 45 shown in FIG. 2. As a result, the end face 44 of the first piston 32 in the fourth hydraulic chamber 45 can be acted upon by the pressure prevailing in the differential pressure chamber 5.

For the rest, the embodiment variant shown in FIG. 2 of a fuel injector with pressure intensifier, which is controlled via a servo valve, corresponds to the embodiment variant already described in connection with FIG. 1.

The mode of operation of the fuel injection system with pressure intensifier shown in FIGS. 1 and 2 is as follows: In the initial state, ie when the second switching valve 24 is closed, the control chamber 29 of the servo valve 23 is acted upon by the pressure prevailing in the high-pressure source 1 (high-pressure storage chamber) via the supply line 22. A closing compressive force acts on the end face 28 of the second piston 33, which is higher than the compressive force acting on the end face 44 of the first piston 32 in the opening direction from the third hydraulic space 42. The piston assembly 32, 33 is thereby placed in its lower position, so that the first sealing seat 36 is closed and the second sealing seat 40 is opened due to the open slide edge. As a result, the differential pressure chamber 5 (rear chamber) of the pressure booster 2 is acted upon by the pressure prevailing in the third hydraulic chamber 42, which corresponds to the pressure prevailing in the high-pressure source 1, via the second hydraulic chamber 38 via the connecting line 43 and the open flow channel 41. Because of this, the pressure intensifier 2 remains deactivated, since the pressure prevailing in the high-pressure source 1 is also present in its working space 5. To ensure high-pressure tightness, a first covering length 37 is formed below the pressure stage.

By activating the second switching valve 24, the control chamber 29 of the servo valve 23 is relieved of pressure in the third return 25 on the low-pressure side, whereby the piston assembly 32, 33 opens. The hydraulic opening force generated in the third hydraulic chamber 42 on the end face 44 of the first piston 32 enables the piston assembly 32, 33 to open quickly and precisely. In the open state, the second sealing seat 40 is closed, whereas the first sealing seat 36 is open. In this case, the differential pressure space 5 (rear space) of the pressure intensifier 2 is connected via the second hydraulic space 38, the opened first sealing seat 36 and the first hydraulic space 34 to the fourth low-pressure-side return 35 branching off from it, so that the pressure intensifier 2 is activated and Compressed fuel flows into the compression chamber 6 via the high-pressure line 8 to the control chamber 11 of the fuel injector 9 and the nozzle chamber 14 thereof.

If the second switching valve 24 is closed again, the piston assembly 32, 33 moves into its starting position due to the hydraulic pressure force acting in the closing direction in the control chamber 29 of the servo valve 23, which acts on the end face 28 of the second piston 33. Due to the hydraulic closing force, a precisely defined closing movement occurs over the entire stroke of the piston assembly 32, 33. To support the closing movement, a spring force can additionally be provided, which, however, is no longer shown in the embodiment variants of the servo valve 23 according to FIGS. 1 and 2. To stabilize the guidance of the piston assembly 32, 33, an integrated flow channel 41 is formed on the first piston 32 of the piston assembly 32, 33. Instead of the 3/2 variant of the servo valve 23 shown in FIGS. 1 and 2, a 2/2 variant can also be used, as well as a 4/2 variant in which the function of the check valve 7 in the piston assembly 32, 33 of the servo valve 23 can be integrated.

In a slight modification of the embodiment variant shown in FIG. 1, the fourth hydraulic chamber 45 is provided in the embodiment variant shown in FIG. 2, in which the compressive force acting on the end face 44 of the first piston 32 in the opening direction prevails. The fourth hydraulic chamber 45 is connected via the connecting line 46 to the differential pressure chamber 5 (rear chamber) of the drain converter 2. According to this embodiment variant, the first phase of the closing movement of the piston assembly 32, 33 can be accelerated.

FIG. 3 shows a variant of a fuel injector, the pressure intensifier assigned to it also being controlled via a servo valve.

Deviating from the booster piston 3 of the pressure booster 2 used in the embodiment variants according to FIGS. 1 and 2, a booster piston 50 with an integrated check valve is provided in the embodiment variant according to FIG. Furthermore, the control chamber 29 of the servo valve 23 is pressurized via a second inlet throttle 26 connecting the working chamber 4 of the pressure intensifier 2 directly to the control chamber 29. This is not in the supply line 22, via which the working chamber 4 of the pressure intensifier 2 as shown in FIG. 3 with the high pressure source 1 (high pressure storage space) is integrated.

The fuel injector 9 according to FIG. 3 corresponds to the fuel injector which has already been described in connection with FIGS. 1 and 2.

The servo valve 23 according to FIG. 3 is designed as a servo-hydraulically assisted valve and comprises a first valve piston part 32, to which a smaller-diameter second piston part 33 is assigned. The valve piston is made in one piece. The servo valve 23 is activated or deactivated by actuating the second switching valve 24. The second switching valve 24 is assigned a third return 25 on the low-pressure side, via which the control chamber 29 of the servo valve 23 can be relieved of pressure by interposing the second discharge throttle 27 in the third return 25 on the low-pressure side. The booster piston 50 of the pressure booster 2 according to the embodiment variant in FIG. 3 comprises a through-channel 51 which connects the working space 4 with the compression space 6 pressure booster 2. The compression space 6 is refilled via the work space 4 via the check valve 7 integrated in the booster piston 50.

In contrast to the embodiment variants shown in FIGS. 1 and 2 with regard to the first hydraulic chamber 34 on the servo valve 23, in the embodiment variant according to FIG. 3 this is not formed in the valve housing 47 of the servo valve 23, but on the piston as a constriction 52.

FIG. 3 shows the switching position of the servo valve 23 in which the pressure intensifier 2 is deactivated. In the control chamber 29, when the second switching valve 24 is in its seat, the pressure level prevailing in the high-pressure source 1 (high-pressure storage chamber) is present via the second inlet throttle 26 branching off from the working chamber 4 and the supply line 22. The pressure force acting on the end face 44 of the first valve piston part 32 pushes it into its upper position, since the closing force acting on the end face 44 is greater than the pressure force acting on the annular pressure stage in the third hydraulic chamber 42 and acting in the opening direction. In this position of the first valve piston part 32, the first sealing seat 36 is closed due to the overlap length 37, whereas the second sealing seat 40 in the housing 47 of the servo valve 23 is open. Because of this, the differential pressure chamber 5 (rear chamber) of the pressure booster 2 is acted upon by the pressure prevailing in the third hydraulic chamber 42 via the open second sealing seat 40 and the second hydraulic chamber 38, so that the pressure booster 2 remains deactivated.

To ensure sufficient high-pressure tightness of the second hydraulic chamber 42 with respect to the fourth hydraulic chamber 45 on the low-pressure side and the fourth low-pressure side return 35 branching off from this, the first overlap length 37 is formed on the second valve piston part 33. Because of the second valve piston part 33, the first overlap length 37 is in the embodiment variant according to FIG

3 compared to the first covering length 37 in the embodiment variants according to FIGS. 1 and 2 significantly reduced.

FIG. 4 shows the fourth state of the switching valve which controls the pressure intensifier of a fuel injector.

Starting from the initial state shown in FIG. 3, when the second switching valve 24 according to FIG. 4 is activated, the control chamber 29 of the servo valve 23 becomes the second one Discharge throttle 27 relieved in the third low-pressure return 25. Due to the decreasing pressure in the control chamber 29, the piston 32 moves with its end face 44 to a stop 30. The opening movement of the first valve piston part 32 and the second valve piston part 33 is supported by the hydraulic opening force generated in the third hydraulic chamber 42. This is connected via the overflow line 39 to the differential pressure chamber 5 (return chamber) of the pressure booster 2, from which a not inconsiderable control volume flows out via the third hydraulic chamber 42, the fourth hydraulic chamber 45 into the fourth low-pressure return 35 when the pressure is released. In the cold state of the servo valve 23 shown in FIG. 4, the second sealing seat 40 is closed, whereas the first sealing seat 36 is open due to the first overlap length 37 extended from the housing 47 of the servo valve 23. The differential pressure chamber 5 (return chamber) of the pressure intensifier 2 is now connected to the fourth low-pressure-side return via the third hydraulic chamber 42 and the opened first sealing seat 36 via the fourth hydraulic chamber 45, so that the intensifier piston 50 with integrated check valve 7 into the compression chamber 6 of the pressure intensifier 2 retracts. As a result, both the control chamber 11 of the fuel injector 9 and, via the nozzle chamber inlet 15, the nozzle chamber 14 of the fuel injector 9 are acted upon by fuel under increased pressure.

When the second switching valve 24 is actuated again, i.e. When the third low-pressure return is closed, pressure builds up in the control chamber 29 of the servo valve 23, so that the first valve piston part 32 and the second valve piston part 33 move back into the starting position shown in FIG. 3. A hydraulic closing force generated in this way achieves a fast, precisely defined closing movement in the servo valve 23 over the entire stroke of the valve piston, including the first valve piston part 32 and the second valve piston part 33. To support the closing movement 23 spring elements can be provided in the control chamber 29 of the servo valve.

Analogous to the design of the second piston 32 according to the embodiment variants in FIGS. 1 and 2, integrated flow channels 41 can be provided on the second valve piston part 33 of the valve piston as shown in FIGS. 3 and 4, which serve to stabilize the piston movement in the servo valve 23.

FIG. 5 shows a further embodiment variant of a servo valve which controls a drain intensifier of a fuel injector.

The variant of the servo valve 23 shown in FIG. 5 is in its initial state, ie in its closed position. The drain translator 2 shown in the embodiment variant according to FIG. 5 corresponds to the execution of the drain translator 3 and 4 with an integrated check valve 7. The fuel injector 9 is constructed identically to the fuel injectors already described in connection with FIGS. 1, 2, 3 and 4. In a departure from the previously described embodiment variants of the servo valve 23 proposed according to the invention, the servo valve 23 comprises a multi-part housing 61 which has a first housing part 62 from which the fourth return line 3 on the low-pressure side branches off and a second housing part 63 which accommodates the one-part valve piston of the servo valve 23. The valve piston 60 comprises a first valve piston part 32 and a valve piston part with a reduced diameter. A further seal 64 is formed on the underside of the first housing part 62 of the multi-part housing 61, opposite the end face 28 of the reduced-diameter valve piston part. The seal 64 can be designed as a flat seat, as a conical seat or as a ball seat. One or more flow channels 41 are arranged on the circumference of the valve piston part with reduced diameter. The overlap length 37 on the outer circumference of the reduced-diameter valve piston 60 is further reduced in comparison with the overlap lengths 37 on the second valve piston part 33, as shown in FIGS. 3 and 4.

In the initial state shown in Figure 5, i.e. In this switching position of the servo valve 23, the pressure level present in the high-pressure source is present in the control chamber 29 of the servo valve 23 via the second inlet throttle 26, the working chamber 4 of the pressure booster 2 and the supply line 22 branching off from the high-pressure source (high-pressure storage space). The second switching valve 24 closes the third return 25 on the low-pressure side. Due to the pressure prevailing in the control chamber 29, a pressure force acting in the closing direction acts on the end face 44 of the first valve piston part 32. This is greater than the draining force acting in the opening direction on the first valve piston part 32 on the annular surface in the third hydraulic chamber 42, so that the first valve piston part 32 is set in the position shown in FIG. 5, sealing the seal 64. In this position of the valve piston 60 of the servo valve 23, the first sealing seat 36 is closed, whereas the second sealing seat 40, designed as a slide seal, is open. Due to the sealing of the fourth hydraulic space 45 by the closed seal 64, when the servo valve 23 is closed, there is no leakage flow into the fourth low-pressure side return 35 become.

The seal 64 can be implemented in a variety of ways, and its design can be represented as a flat seat, a conical seat or a ball seat. This is particularly advantageous Execution of the seal 64 as a flat seat in connection with a multi-part housing 61 of the servo valve 23. If the seal 64 is formed in particular as a flat seat in a separate housing part 62, a possibly occurring axial misalignment between the valve piston 60 of the servo valve 23 and the housing 62 can be compensated for. With the design of the servo valve 23 shown in FIG. 5, a high closing force, which improves the sealing effect, is applied to the valve piston 60 of the servo valve 23, as a result of which a very high surface pressure and thus a good sealing effect result on the seal 64 when it is designed as a flat seat. In the idle state of the servo valve 23 shown in FIG. 5, the differential pressure chamber 5 (Rücl aum) of the pressure booster 2 is above the open sealing edge 40 and the second hydraulic chamber 38 formed in the second housing part 63, with the third hydraulic chamber 42 interposed with that in the high-pressure source 1 ( High-pressure storage space) prevailing pressure in connection. The pressure intensifier 2 is thus deactivated, since the same pressure prevails both in the work space 4 and in the differential pressure space 5 (rear space).

When the second switching valve 24 is activated, the control chamber 29 of the servo valve 23 is relieved of pressure.

FIG. 6 shows the servo valve according to the variant in FIG. 5 when actuated by the second switching valve 24.

Due to a pressure relief of the control chamber 29 of the servo valve 23, fuel flows via the second switching valve 24 into the third low-pressure side return 25. The valve piston 60 of the servo valve 23 moves towards a stop 30 formed in the control chamber 29 of the servo valve 23. The end face 44 of the valve piston 60 bears against this stop 30, as shown in FIG. The hydraulic opening force generated in the third hydraulic space 42, due to the control volume flowing over from the differential pressure space 5 via the overflow line 39, results in a rapid, exact opening. When the valve piston 60 opens, the seal 64 is first opened and the sealing edge 40 is closed. Only then does the first sealing seat 36, which is designed as a slide seal, open. This can prevent a short-circuit leakage flow from the second hydraulic chamber 38 from entering the fourth return 35 on the low-pressure side. The differential pressure chamber 5 of the pressure converter 2 is now via the third hydraulic chamber 42, the open slide seal 36, the open seal 64 and a further hydraulic chamber 65 formed in the first housing part 62 with the fourth low-pressure side return 35 connected. The pressure booster 2 is thus activated and compresses the fuel volume contained in the compression space 6.

When the second switching valve 24 is actuated again and the associated re-filling of the control chamber 29 of the servo valve 23, the valve piston 60 of the servo valve 23 moves into its starting position as shown in FIG. 5 by the hydraulic pressure force building up in the control chamber 29 The hydraulic closing force in the control chamber 29 of the servo valve 23 ensures an exactly defined closing movement over the entire stroke range of the valve piston 60. To support the closing movement, spring elements integrated in the control chamber 29 can also be used, but are not shown in the illustrations according to FIGS. 5 and 6. When the servo valve 23 is closed, the first sealing seat (slide seal 36) is first closed. By closing the slide seal 36, the differential pressure chamber 5 (rear chamber) of the drain converter 2 is decoupled from the fourth low-pressure side 35. Only after a further closing stroke of the valve piston 60 and thus after a _{delay time} t _l3 does the sealing edge 40 open, so that the pressure intensifier 2 is only then completely deactivated. As valve piston 60 moves further in the direction of seal 64, it closes. As a result of the delay time X _\ , after a main injection has been carried out, a pressure cushion remains in the nozzle chamber 14 of the fuel injector for a short time, which can be used for post-injection under high pressure. This switching sequence of opening or closing the sealing points 36, 40, 64 can avoid an overlap of opening cross sections, ie no phase with simultaneous opening of two flow cross sections occurs during the movement of the valve piston.

The reduced-diameter piston part of the valve piston 60 according to the representations in FIGS. 5 and 6 comprises one or more integrated flow channels 41 to stabilize the piston movement in the guide area. The return lines 19, 21, 25, 35 can be separated from one another instead of in FIGS. 1 to 6 trained returns can also be partially or completely combined and connected to a return system common to all returns. LIST OF REFERENCE NUMERALS Hochdrackquelle (Hochdrackspeicherraum) pressure booster booster piston working space Differenzdrackraum (return chamber) compression chamber check valve high-pressure passage fuel injector the injection valve member control chamber first inlet throttle first outlet throttle nozzle chamber nozzle chamber inlet injection port pressure stage first shift valve first niederdrackseitiger recoil spring piece of second low-pressure side return supply line servo valve second switching valve third low-pressure side return second inlet throttle second outlet throttle piston end face control chamber servovalve , Stop valve spring, first valve piston part, second valve piston part, first hydraulic chamber, fourth return on the low-pressure side, slide valve seal, first overlap length (h _ϊ ) second hydraulic chamber overflow line sealing edge integrated flow channel third hydraulic chamber connecting line differential pressure chamber second hydraulic chamber front side fourth hydraulic chamber connecting line differential pressure chamber - fourth hydraulic chamber servo valve housing

Booster piston with integrated check valve, through-passage constriction

Valve piston multi-part housing first housing part second housing part Sealing further hydraulic space

Claims

claims

1. Servo valve for actuating a pressure intensifier (2), which is assigned to a fuel injector (9), the pressure intensifier (2) having a work space (4) which is separated from a differential pressure space (5) by a intensifier piston (3, 50) , and the pressure change in the differential pressure chamber (5) of the pressure booster (2) takes place via the servo valve (23) to which a switching valve (24) activating it is assigned, characterized in that in a valve housing (47; 61, 62, 63) Servo valve (23), the control chamber (29) of which can be connected both to a high-pressure source (1) and can be relieved of pressure in a low-pressure outlet (25) and for generating a fast closing movement on the valve piston (32; 33, 60) between the control chamber (29) and a hydraulic chamber (42) a pressure stage (44, 28) acting in the closing direction of the valve piston (32, 33, 60) is formed and on the valve piston (32, 33, 60) control edges (36, 40) without common oil opening phase are trained ..

2. Servo valve according spoke 1, characterized in that the valve piston has a first valve piston part (32) and a second valve piston part (33) with reduced diameter.

3. Servo valve according to claim 2, characterized in that an overlap length (37) forming a slide seal (36) is formed on the valve piston part (32) which is reduced in diameter.

4. Servo valve according to claim 2, characterized in that one or more flow channels (41) are formed on the reduced-diameter valve piston part (33) of the valve piston (32, 33, 60).

5. Servo valve according to claim 2, characterized in that the separation point of the first valve piston part (32) and the reduced diameter second valve piston part (33) is in a low-pressure-side space (34), end faces (28, 44) of the valve piston parts ( 32, 33) are subjected to high pressure.

6. Servo valve according spoke 1, characterized in that a starting from the control chamber (29) guide portion in the servo valve housing (47) opens into a high-pressure second hydraulic chamber (29, 38).

7. Servo valve according to claim 6, characterized in that the guide section of the first valve piston part (32) in the servo valve housing (47) is formed without a valve pocket.

8. Servo valve according to claim 6, characterized in that a further seal (64) is formed on the valve piston (60), which cooperates with a housing part (62) of a multi-part valve housing (61).

9. Servo valve according to claim 8, characterized in that the further seal (64) is designed as a flat seat.

10. Servo valve according to claim 6, characterized in that integrated flow ducts (41) allowing fuel to flow out are formed on the valve piston (60) above a cover length (37) with a second housing part (63) of the multi-part housing (61) ,

11. Servo valve according spoke 1, characterized in that an effective in the opening direction of the servo valve piston (32, 33) pressure surface (44) is acted upon by the pressure prevailing in the differential pressure chamber (5).

12. Servo valve according spoke 1, characterized in that when the servo valve is deactivated, the low-pressure side (5) is sealed against the high-pressure side (38, 39) by a guide section (37) on the valve piston (32, 33, 60).