HIGH-PRESSURE PUMP

Abstract

A high-pressure pump (1), which serves in particular as a radial or inline piston pump for fuel injection systems of air-compressing, auto-ignition internal combustion engines, comprises a cylinder head (2) and a pump assembly (6). Here, the cylinder head (2) has a cylinder bore (4) in which a pump piston (5) of the pump assembly (6) is guided. Here, the pump piston (5) delimits, in the cylinder bore (4), a pump working chamber (12). Also provided is an inlet valve (20) which is integrated into the cylinder head (2) and via which fuel can be conducted into the pump working chamber (12). Metering of the fuel conducted into the pump working chamber (12) can be achieved by actuation of the inlet valve (20). Here, full charging of the pump working chamber (12) may take place. It is however also possible for partial charging of the pump working chamber (12) to be achieved by means of suitable actuation of the inlet valve (20).

Description

BACKGROUND OF THE INVENTION

The invention concerns a high-pressure pump, in particular a radial or in-line piston pump. The invention concerns in particular the field of fuel pumps for fuel injection systems of air-compressing, auto-ignition internal combustion engines. The high-pressure pump can however also be used as a piston pump to deliver other suitable fluids.

DE 195 15 191 A1 discloses a high-pressure fuel pump. The high-pressure fuel pump has a cylinder, the upper part of which lies open towards the outside of the head cover which is part of the engine housing. The remaining segment of the high-pressure fuel pump is accommodated in a housing hole in the head cover. A pump cam is mounted on a valve camshaft to drive an intake/exhaust valve and drives the high-pressure fuel pump. As the time behavior with which the pressurized fuel is expelled is controlled by activation of a solenoid valve, the accuracy with which the fuel delivery is controlled is also improved.

The high-pressure fuel pump disclosed in DE 195 15 191 A1 is a pump choked on the suction side which has several disadvantages. The disadvantages are high noise, poor controllability and the occurrence of mechanical vibrations due to cavitations occurring in the supply lines to the inlet valves. Pressure waves between a feed metering unit and the suction valve have an unfavorable effect on the function.

SUMMARY OF THE INVENTION

The high-pressure pump according to the invention has the advantage that an improved design is achieved in which in particular a metered fuel feed and compact design are possible. In particular no feed metering unit or similar is required, leading to a substantial cost reduction in production.

In contrast to high-pressure pumps with suction-side volume flow control by means of a feed metering unit in combination with spring-loaded intake valves, which have the drawback that at a high pump speed, constant delivery cannot be guaranteed and that pressure fluctuations in the low pressure circuit can lead to noise, advantageously a cost reduction can be achieved by the omission of a feed metering unit, even at high pump speeds a constant delivery can be made possible and noise reduction can be achieved by the avoidance of pressure fluctuations and possible cavitation in the low pressure circuit.

In a conventional design, in particular in multi-piston pumps with three or more pistons, the suction phases overlap. Pressure fluctuations then lead to particularly great differences in the quantity delivered. This can advantageously be avoided. Here it is possible to exclude such differences in the pre-stored quantity.

In particular a great cost advantage is achieved with a high-pressure pump designed as a single piston pump. Even when designed as a two-piston pump with a further actuator, the absence of bores in the housing of the high-pressure pump partially compensates for extra costs. One essential advantage of direct control is the expansion of the pump rotation speed range and hence an improvement in the efficiency of the high-pressure pump.

Also integration of the inlet valve into the cylinder head allows a very small construction size. This also applies for very high pressures, for example of 300 MPa (3000 bar) as conceivable for application on trucks.

Advantageously the inlet valve is formed as a magnetically controllable inlet valve. Furthermore it is advantageous that the inlet valve is fixed to the cylinder head by means of a screw plug screwed into the cylinder head, and that the screw plug is formed of a ferromagnetic material. As a result the screw plug can serve as a magnetic conductor which improves the efficiency of the magnetic circuit and allows a high magnetic force.

Also it is advantageous for a magnet coil to be provided, that the inlet valve can be controlled via current flowing through the magnet coil, and that the magnet coil can be cooled by fuel that can be transferred via the inlet valve to the pump working chamber. Thus cooling of the magnet coil and further elements of the magnetic circuit can be achieved by flushing with the fuel.

It is also advantageous that the inlet valve has a valve body and a valve tappet co-operating with the valve body to form a seal seat, wherein the valve tappet lies on the cylinder head, wherein a magnetically activatable solenoid plunger is provided and wherein the solenoid plunger carries the valve tappet with it on mechanical activation to open the seal seat formed between the valve body and the valve tappet. Thereby the magnetic force to activate the inlet valve can be generated via the solenoid plunger, wherein the screw plug advantageously serves as a magnetic conductor. The inlet valve is here preferably closed when the magnet coil is switched without current. If current flows through the magnet coil of the magnet and the pump piston is for example at the top dead center, then the inlet valve opens. On full filling, the inlet valve is preferably open until the bottom dead center of the pump piston. It is furthermore advantageous here that an adjustment shim is provided which serves to specify a working air gap and a residual air gap for the solenoid plunger. This allows a modular design wherein by fitting of a suitable adjustment shim, adaptation is possible to the respective application of the high-pressure pump. This expands the application range of the high-pressure pump, wherein simple adaptation and largely identical design of the high-pressure pump are possible.

It is also advantageous that a controller is provided which controls the inlet valve as a function of movement of the pump piston of the pump assembly. Firstly it is advantageous that the controller, to reduce the filling of the pump working chamber of the pump assembly, shortens the control time at its end so that the inlet valve is closed before the pump piston reaches the bottom dead center, or extends the control time at its end so that the inlet valve is closed after the pump piston reaches a bottom dead center. Thus the control time can be reduced so that the inlet valve is closed again before the pump piston reaches the bottom dead center, which reduces the quantity of fuel flowing into the pump working chamber. This can alternatively also be achieved in that the injector valve is only closed after the reaching of the bottom dead center, whereby the fuel transferred to the pump working chamber is returned partly via the inlet valve in the opposite direction by the movement of the pump piston. In the first case the pressure fluctuations on the low pressure side are reduced. In the second case preferably no cavities occur in the working cylinder. The advantageous variant may be selected specially depending on the application. A further possibility is that the control time is shortened at its start so that the inlet valve is opened only after the pump piston reaches a top dead center. Thus the inlet valve is not opened immediately after the top dead center, so that the quantity of fuel flowing into the pump working chamber is also reduced. Here also a suitable combination of control methods can be performed by the controller. For example the control time can be shortened both at its start and at its end. Thus advantageously partial fillings of the pump working chamber can be achieved. Furthermore pressure fluctuations can be influenced positively with regard to amplitude and frequency by one or more chokes connected before the inlet valve. Also the quantity regulation can be positively influenced. In this way the noise behavior, which can be unfavorably affected by pressure fluctuations in the low pressure, can be improved.

The inlet valve is preferably fitted with a closing spring with a high spring pretension to achieve a high closing dynamic.

BRIEF DESCRIPTION OF THE DRAWING

Preferred embodiment examples of the invention are described in more detail in the description below with reference to the enclosed drawing. This shows:

FIG. 1 a high-pressure pump in an extract, schematic, axial cross section view corresponding to one embodiment example of the invention.

DETAILED DESCRIPTION

FIG. 1 shows a high-pressure pump 1 in an extract, schematic, axial cross section view corresponding to one embodiment example of the invention. The high-pressure pump 1 can in particular be designed as a radial or in-line piston pump. The high-pressure pump 1 is particularly suitable as a fuel pump for fuel injection systems of air-compressing, auto-ignition internal combustion engines. A preferred use of the high-pressure pump 1 lies in a fuel injection system with a fuel distribution rail which stores diesel fuel under high pressure. The high-pressure pump 1 according to the invention is however also suitable for other applications. In particular the high-pressure pump can be used as a piston pump to deliver suitable fluids, in particular fluids other than fuel.

The high-pressure pump 1 has a pump housing which is mounted on a cylinder head 2. The cylinder head 2 has a shoulder 3 which protrudes into a bore in the pump housing. Here in the shoulder 3 is formed a cylinder bore 4 in which a pump piston 5 of a pump assembly 6 is guided along an axis 7.

The high-pressure pump 1 also has a drive shaft 8 on which is provided a cam 9. The cam 9 can here also be designed as a multiple cam or as an eccentric segment of the drive shaft 8. In operation the drive shaft 8 with the cam 9 rotates about a rotary axis 10. Between the pump piston 5 of the pump assembly 6 and the cam 9 is an active connection 11 which is illustrated by the double arrow 11. For example via a roller shoe and a roller mounted in the roller shoe, an actuation force can be transferred by the cam 9 to the pump piston 5. A return of the pump piston 5 can take place via a suitable tappet spring.

Thus the pump assembly 6 can be driven by the cam 9 of the drive shaft 8. Depending on the design of the high-pressure pump 1, further pump assemblies can also be driven by the cam 9. Also on the drive shaft 8 can be provided further cams which serve to drive further pump assemblies. Depending on the design, a high-pressure pump 1 can thus be constructed as a radial or in-line piston pump.

The pump piston 5 in the cylinder bore 4 delimits a pump working chamber 12. A supply channel 13 serves to supply fuel which is delivered therein by a pre-delivery pump. In the supply channel 13 are provided a first choke 14 and a second choke 15. The supply channel 13 leads into a low pressure chamber 16 which is formed by a recess 17 in the cylinder head 2.

The high-pressure pump 1 has an inlet valve 20. The low pressure chamber 16 is here part of the inlet valve 20. The inlet valve 20 is integrated in the cylinder head 2. The inlet valve 20 is arranged in the recess 17 of the cylinder head 2. The recess 17 is here closed by a screw plug 21. Thus the low pressure chamber 16 is closed to the environment. The screw plug 21 acts via a valve part 22 on a valve body 23. The screw plug 21 is screwed into the cylinder head 2 and through this presses the valve body 23 against a contact surface 24 formed on the cylinder head 2. The screw plug 21, valve part 22 and valve body 23 of the inlet valve 20 are thus fixed in a stationary manner. Also the screw plug 21 and valve part 22 are preferably formed of ferromagnetic material.

In the valve body 23 is guided a valve tappet 25. Here the valve tappet 25 co-operates with a valve seat surface 26 formed on the valve body 23 to form a seal seat. A valve spring 27 presses the valve tappet 25 against the valve seat surface 26. The valve spring 27 here acts via a valve element 28 and an adjustment shim 29 on a rotor 30. The rotor 30 is formed as a solenoid plunger 30. The solenoid plunger 30 is connected with the valve tappet 25. Thus the valve tappet 25 is pressurized by the pretension of the valve spring 27. The valve tappet 25, valve element 28, adjustment shim 29 and solenoid plunger 30 of the inlet valve 20 are mobile elements which are moved to open the inlet valve 20 on control of the inlet valve 20.

The inlet valve 20 also has a magnet 31 with a magnet coil 32. The magnet coil 32 is electrically connected via electrically conductive contact pins 33, 34 with pins 35, 36 of a plug 37. The plug 37 here allows connection with a control unit 38. The control unit 38 in this embodiment example serves as a controller 38. The controller 38 can also be integrated in a central control unit. The control unit 38 is connected with a rotary angle sensor 39 which detects the momentary rotation angle of the drive shaft 8 and emits this to the control unit 38. Via the rotary angle detected, there is a direct connection to the momentary position of the pump piston 5. It can in particular be detected thus whether the pump piston 5 is at a top dead center at which the pump piston undergoes a maximum stroke and the pump working chamber 12 has minimum volume. Accordingly it can be detected whether the pump piston 5 is at a bottom dead center at which the pump piston 5 has minimum stroke and the volume of the pump working chamber 12 is maximum.

Current flowing through the magnet coil 32 generates a magnetic flux. This magnetic flux is emitted by the magnet 31 wherein an amplification is possible via the ferromagnetic screw plug 21. The magnetic flux also runs via the valve part 22, the solenoid plunger 30 and where applicable further ferromagnetic elements back to the screw plug 21. Between the solenoid plunger 30 and the valve part 22 there is a gap 40. The gap 40 firstly allows displaceability of the solenoid plunger 30 and thus an adjustment of the valve tappet 25 to activate the inlet valve 20. Secondly at least one rest air gap remains as gap 40 in order to avoid, in activated state, a so-called magnetic adhesion effect of the solenoid plunger 30 on the valve part 22. In particular when the power to the magnet coil 32 is switched off, the force of the valve spring 27 can initiate a closure of the inlet valve 20 largely without distortion. The maximum size of the gap 40 is specified by the sum of the desired working air gap and the residual air gap. Adjustment of the residual air gap and working air gap is possible by suitable selection of valve element 28 and adjustment shim 29. In particular the thickness of the adjustment shim 29 can pre-specify the desired working air gap. The thickness of the adjustment shim 29 thus specifies the stroke of the valve tappet 25. With unchanged geometry in the region of the valve seat surface 26, thus the opening cross section at the valve seat surface 26 can be changed and hence also the possible throughflow into the pump working chamber 12 set when the seal seat is opened. Thus the inlet valve 20 can be adapted in relation to the application concerned.

By actuating inlet valve 20, fuel can thus be guided from the low-pressure chamber 16 to the pump working chamber 12. Activation of the inlet valve 20 here takes place during a suction stroke of the pump piston 5. During the delivery stroke of the pump piston 5, the inlet valve 20 is preferably closed. Thus fuel under high pressure is delivered via outlet valve 41—which can be designed as a directional or non-return valve 41—into a high-pressure line 42. The high-pressure line 42 is for example connected with a fuel distribution rail.

If the inlet valve 20 is opened at approximately the top dead center of a pump piston 5 and closed at the bottom dead center of pump piston 5, a full filling of the pump working chamber 12 can be achieved. However the inlet valve 20 can be controlled by the controller 38 irrespective of the stroke or momentary position of pump piston 5 during the suction phase. This allows also a partial filling of the pump working chamber 12. There are several possibilities for this which can be combined where applicable.

One possibility is to reduce the control time of the inlet valve 20 so that the inlet valve 20 is closed again before the pump piston 5 reaches the bottom dead center. Alternatively the control time can also be extended beyond the reaching of the bottom dead center. The inlet valve 20 is thus only closed after pump piston 5 reaches the bottom dead center so that part of the fuel is delivered back from the pump working chamber 12 during the stroke of the pump piston 5 in the direction back through the inlet valve 20. The other part of the fuel is delivered then via the high-pressure line 42. The quantity of fuel delivered via the high-pressure line 42 per pump stroke is thus reduced.

It should be noted that there is no shut-off control of fuel to a tank or similar. Also in this way where applicable noise behavior can be improved by damping pressure pulses. Adjustment is possible via chokes 14, 15.

A further possibility of achieving a partial filling is that the inlet valve is not opened immediately after pump piston 5 reaches the top dead center. This achieves a certain idle stroke of the pump piston 5 so that the entirety of fuel flowing into the pump working chamber 12 via the opening cross section of the opened seat seal is thus reduced.

Thus advantageously through one or more chokes 14, 15 or damping volumes connected before the intake valve, the pressure fluctuations can be reduced with regard to amplitude and frequency and the quantity control can be reduced. The chokes here allow a large partial reflection and slight damping of pressure and attenuation waves. Damping volumes allow a lower partial reflection and greater damping of the pressure and attenuation waves. This is dependent on the geometric design of the respective damping volume. By opening and closing the inlet valve 20 or where applicable several inlet valves designed corresponding to the inlet valve 20, pressure and attenuation waves occur which run from the intake valves to a delivery pump, in particular an electric fuel pump, and are reflected there. The reflected waves can inter alia make contact again on an opening process of the inlet valve 20 and thus further influence the filled mass in the pump working chamber, which can lead to delivery fluctuations of the high-pressure pump. By use of damping volumes and chokes 14, 15 in the supply channel 13 and by matching these, such pressure waves can be reduced so far that a constant delivery of the high-pressure pump 1 is guaranteed within a certain tolerance range. The design and dimensioning here depend on the area of application of the high-pressure pump 1 and the connection to the pre-delivery pump.

Advantageously thus an intake valve 20 can be produced which is closed in the unpowered state. This inlet valve 20 is integrated in the cylinder head 2. The solenoid plunger principle can be utilized here so that rapid opening and closing of the intake valve 20 can be achieved. Furthermore the suction choking can be shifted into the working cylinder with deliberate utilization of an air exhalation. The necessary dynamic can be guaranteed by one or more connecting bores. A sufficiently high closing dynamic can be achieved via a correspondingly high spring pretension of valve spring 27. The cooling of the magnet 31 with the magnet coil 32 can be achieved by flushing with fuel.

The invention is not restricted to the embodiment examples described.

Claims

1. A high-pressure pump (1) for fuel injection systems of air-compressing, auto-ignition internal combustion engines, with at least one cylinder head (2) and a pump assembly (6), wherein the cylinder head (2) has a cylinder bore (4) in which is guided a pump piston (5) of the pump assembly (6), wherein the pump piston (5) in the cylinder bore (4) delimits a pump working chamber (12), wherein an inlet valve (20) integrated in the cylinder head (2) is provided via which fuel can be transferred to the pump working chamber (12) and whereby, by control of the inlet valve (20), a metered feed of the fuel guided into the pump working chamber (12) is possible.

2. The high-pressure pump as claimed in claim 1, characterized in that the inlet valve (20) is formed as a magnetically controllable inlet valve (20).

3. The high-pressure pump as claimed in claim 2, characterized in that the inlet valve (20) is fixed to the cylinder head (2) by means of a screw plug (21) screwed into the cylinder head (2) and that the screw plug (21) is made of a ferromagnetic material.

4. The high-pressure pump as claimed in claim 1, characterized in that a magnet coil (32) is provided, that current flowing through the magnet coil (32) allows control of the inlet valve (20), and that the magnet coil (32) can be cooled by fuel which can be transferred via the inlet valve (20) to the pump working chamber (12).

5. The high-pressure pump as claimed in claim 1, characterized in that the inlet valve (20) comprises a valve body (23) and a valve tappet (25) co-operating with the valve body (23) to form a seal seat, wherein the valve tappet (25) lies against the cylinder head (2), wherein a magnetically activatable solenoid plunger (30) is provided and wherein the solenoid plunger (30) carries the valve tappet (25) with it on magnetic actuation to open the seal seat formed between the valve body (23) and the valve tappet (25).

6. The high-pressure pump as claimed in claim 5, characterized in that an adjustment shim (29) is provided which serves to specify a working air gap for the solenoid plunger (30).

7. The high-pressure pump as claimed in claim 1, characterized in that a controller (38) is provided which controls the inlet valve (20) as a function of movement of the pump piston (5) of the pump assembly (6).

8. The high-pressure pump as claimed in claim 7, characterized in that to reduce a filling of the pump working chamber (12) of the pump assembly (6), the controller (38) at least one of: a) shortens the control time at its end so that the inlet valve (20) is closed before the pump piston (5) reaches a bottom dead center, or extends the control time at its end so that the inlet valve (20) is closed after the pump piston (5) reaches a bottom dead center, andb) shortens the control time at its start so that the inlet valve (20) is opened after the pump piston (5) reaches a top dead center.

9. The high-pressure pump as claimed in claim 1, characterized in that the inlet valve (20) has a low pressure chamber (16) which is formed in a recess (17) of the cylinder head (2) in which the inlet valve (20) is arranged and in that at least one of the low pressure chamber (16) is closed by a screw plug (21) of the inlet valve (20), and that a supply channel (13) leading to the low pressure chamber (16) is provided and that in the supply channel (13) is arranged at least one of at least one choke (14, 15) and at least one damping volume.

10. The high-pressure pump as claimed in claim 1, characterized in that the inlet valve (20) has a closing spring (27) and that the closing spring (27) has a high spring pretension.

11. The high-pressure pump as claimed in claim 1, characterized in that the high-pressure pump is one of a radial and an in-line piston pump.