HRP950445A2 - A fuel injector for an internal combustion engine - Google Patents

Description

This invention relates to a fuel injector for an internal combustion engine, particularly for a large two-stroke diesel engine, having an external housing for mounting in the cylinder head and a completely free fuel passage opening into the nozzle, and having a movable valve slide longitudinally in the slide guide and goes up in the direction towards its abutment of the valve by the action of the preloaded compression spring and in the opposite direction by the fuel pressure in the fuel passage, the first piston, axially movable in the housing, is located at one end of the compression spring and has a first surface which faces away from the spring and, together with the stationary part, defines the first pressure chamber which is connected to the fuel passage through the channel, said pressure spring adjusts the first piston in the direction towards the end position with the minimum volume of fuel in the first pressure chamber.

In such an injector described in Danish patent no. 152 619, which corresponds to Japanese patent no. 1851989, the shape of the fuel passage channel to the first pressure chamber is designed as a throttle channel so that the chamber pressure follows the instantaneous fuel passage pressure with a delay. When the fuel pressure increases during the injection period, the pressure also increases in the pressure chamber, so that the first piston is pushed against the closing spring and increases the force with which the spring acts on the valve slide in the direction towards its abutment, which increases the closing pressure of the injector.

This helps to solve the problem that the closing pressure, i.e. the pressure at which the valve slide moves towards its seat at the end of the delivery period, is usually less than the opening pressure, i.e. the pressure in the fuel passage at which the valve slide is lifted from its valve seat at the beginning of the Delivery period. The lower closing pressure is due to the fact that in the closed position of the injector, the fuel pressure acts on the effective surface of the valve slide which is less than when the valve slide is in its open position, when the pressure also acts on the surface of the slide under the adjacent surface.

In the injector described in the Danish patent, the first piston performs constant adjustment movements during and immediately after the delivery period of the injector, which can result in non-negligible wear of the piston guide surface with systematic leakage from the pressure chamber. At the end of each delivery period, the pressure spring pushes the first piston back to its end position with the minimum volume of fuel in the pressure chamber, so that the injector opening pressure is not affected by the hydraulic load on the pressure spring.

It is known that the compression pressure in the engine cylinder depends on the load, so that at full loads the pressure is significantly higher than at low loads. At full load, for example, the compaction pressure may be at approximately 120 bar, while at idle the compaction pressure is at approximately 40 bar.

When the valve slide is in its closed position, the pressure in the engine cylinder spreads through the nozzle holes and further along the area of the slide under the seating surface, i.e. the section of the slide located on the nozzle side of the valve seat. Because of this, the actual condensing pressure acts on the valve slide with a force in the opening direction. The condensation pressure increasing with the engine load thus leads to a drop in the opening pressure of the known injector at increased engine loads. In a typical fuel injector for a large two-stroke diesel engine, the opening pressure may, for example, drop from 400 bar at idle to 325 bar at full engine load. The lower opening pressure at full load does not promote fuel atomization at the beginning of the injection period.

At low engine loads, the fuel pressure in the injector is determined by its opening pressure, because the fuel pumps deliver such a small amount of fuel that the flow resistance in the injector does not affect the fuel pressure. On the contrary, the quantities of fuel delivered by the pumps at higher engine loads are greater than the flow resistance in the nozzle, they become decisive for the fuel pressure in the injector, i.e. the fuel pressure in this case is significantly higher than the nozzle opening pressure.

The opening pressure in known injectors is determined by overloading the compression spring. The production of springs is subject to certain production tolerances, the consequence of which is the fact that fuel injectors in an internal combustion engine are not necessarily all exposed to exactly the same opening pressure. Changes in injector opening pressures often become more pronounced after a long period of engine operation, because the springs are supplied during operation, i.e. lose some of their spring force. Therefore, there are some time-dependent changes in injector opening pressures, necessitating inspection and re-tensioning of the compression springs at regular intervals to maintain satisfactory engine operation. These jobs are a waste of time and are undesirable.

The object of the invention is to provide a fuel injector that has an increased opening pressure at increased engine loads, and that requires less maintenance.

With this aim in mind, the fuel injector according to the invention is characterized in that the force transmission spring is connected to a second piston having a second surface facing away from the spring and forming an end wall in the second pressure chamber, so that when moving in the direction away from said end the position of the first piston opens an inflow connection between the first and second pressure chambers, so that the second pressure chamber has a larger actual cross-sectional area than the first chamber, and that a restricted flow passage connects the second pressure chamber to the channel.

At increased engine loads, as mentioned above, pressure increases in the fuel passage as a result of resistance to flow in the injector. Therefore, the pressure in the first pressure chamber also increases, which causes the first piston to move so that the flow connection between the first and second pressure chambers is open, and a quantity of fuel flows into the second pressure chamber. Since this pressure chamber has a larger real cross-sectional area than the first chamber, the build-up of pressure in the second chamber causes the first piston to return to its end position with a minimum fuel volume, and at the same time the fuel volume in the second pressure chamber is closed, because the first piston breaks the flow connection between chambers. Since the second force-transmitting piston is connected to the spring, the last one will have a shortened chamber filling step, thus increasing the spring force. Regardless of the small amount of fuel that has leaked out, the fuel is held closed in the pressure chamber until the fuel injector is activated to re-inject the fuel, thus maintaining an increased spring force, which allows the fuel injector to have both a higher opening pressure and a higher closing pressure. .

If the peak pressure in the fuel passage during the next injection period is higher due to an increase in engine load, the fuel pressure in the first pressure chamber will exert a force on the first piston that is greater than the opposing spring force, causing the first piston to move so that the flow connection is open between the chambers, until the build-up of pressure in the second pressure chamber creates an increased spring force slightly greater than the fuel pressure at the actual cross-sectional area of the first pressure chamber. The increase in spring force will then again return the first piston to its end position which breaks the flow connection between the chambers.

A limited discharge passage ensures continuous discharge of a small amount of fuel from the second pressure chamber. This ensures that the spring load is also reduced when the engine load drops and thus the peak pressure in the fuel passage. If the load does not drop, the amount of fuel discharged will now be displaced by the new fuel at the next injection period because the discharge of fuel provides a small pressure drop in the second pressure chamber so that the first piston can open the inlet port again.

The opening pressure of the fuel injector at a specific load in the upper load range of the engine depends on the actual cross-sectional area of the chamber. As such a surface can be made with very tight tolerances, all fuel injectors in the engine will adjust themselves with the same opening and closing pressures, because the different production-related changes in the spring characteristics of the compression springs and their different elastic deflection during the period of operation will be equalized by compressing the springs all while they give the same spring force which is equal to the highest fuel pressure in the injector. This equalization is done automatically while the engine is running, and it makes redundant a significant part of the need for periodic manual adjustment of the injector.

It is possible to construct an inlet connection with a side opening which can be covered or uncovered with the first piston in its displacement, but preferably, the first piston has an abutment portion that closes the inflow connection between the first and second pressure chambers by contact with a corresponding abutment portion on the stationary part. Such mating parts have proven very reliable in fuel injectors over many years, and they provide well-defined pressure differentials capable of closing and withstand large pressure differentials.

The drain passage may conveniently be of such limited size that the drain volume at full engine load during the engine cycle is in the range of one-half to one-twentieth of the entire volume in the second pressure chamber. If the discharge amount becomes more than half, it will be difficult, especially at high engine loads, to achieve the desired increase in opening pressure, and in addition, the displacements of the piston will become larger and more frequent, because the second pressure chamber will have to be refilled in each injection period, also and when the engine load is constant. If the discharged volume is less than one-twentieth, the opening pressure will drop unfavorably slowly with a sudden reduction in engine load. Said discharge ratios correspond to full engine load.

In one embodiment, the effective cross-sectional area of the first pressure chamber may be smaller than the opening area of the valve slide. The consequence of this is that the first piston will remain in the upper end position at low engine loads when the fuel pressure is determined by the injector opening pressure produced only by the mechanical overload of the spring. Only at increased engine load when the fuel pressure increases, the pressure on the effective cross-sectional area of the first chamber will result in a force that can overcome the spring force and move the first piston from its end position.

First of all, the second pressure chamber has an effective cross-sectional area that is several times larger than that of the first chamber. The consequence of this is that the pressure in the second pressure chamber is a corresponding number of times lower than the pressure in the first chamber, when the force of the second piston acts on the spring and thus on the first piston equalizes the oppositely directed force of the fuel on the first piston. The large effective cross-sectional area of the second chamber thus results in the closure of the second chamber at a favorable low pressure in the chamber, which results in a relatively small pressure drop across the discharge passage with a consequent small discharge of fuel from the second pressure chamber. The large area of the second chamber also contributes to the advantage that the chamber is filled with a large volume of fuel at a certain displacement of the second piston and corresponding spring compression. Both of these circumstances contribute to the fact that the spring force changes only slowly while the injector is in its closed position between the two injection periods.

It is possible to place both pistons at the end of the spring closest to the nozzle, because the part that is stationary in relation to the first piston is then created by the valve slide. The result of such a construction is that the pistons participate in adjusting the movement of the valve slide. In this case, the pistons will act as an increase in the mass of the slide, which will give the injector slower movement adjustments. As this is normally considered a disadvantage, the first piston can alternatively be formed at the opposite end of the spring. This results in the disadvantage that the first inlet joint between the two pressure chambers becomes elongated and relatively difficult to manufacture. In a preferred embodiment, which avoids both disadvantages and is simple to manufacture, the fuel injector is designed so that, in a manner known per se, a compression spring is mounted between two spring guides which are longitudinally movable on a centrally pressed part, which is stationary in the housing, that the second piston is formed in the upper spring guide which is located opposite the nozzle and has a lower tubular wall that pressure-sealingly surrounds the inserted part, and an upper tubular wall that has a larger inner diameter than the lower wall and pressure-sealingly surrounds the first piston, and an intermediate the part that interconnects the walls, the upper surface of which forms the second surface, that the first piston is annular and installed between the inserted part and the upper wall of the second piston and has a smaller inner ring of the upper surface that forms the first surface, which merges internally into the adjacent part, and that is the corresponding adjacent part of the inserted part facing downwards and is located between the passage channel towards the pro the fuel line and the lower surface of the reduced diameter that forms the inlet junction between the two chambers.

The discharge passage can be formed as an independent part, for example in the form of a small hole through the second piston into the second chamber, but primarily, the discharge passage consists of pressure-sealing annular gaps between the two walls of the second piston, i.e. between the first piston and the inserted part, which ring the gaps are difficult to make completely pressure-tight, as things stand. The amount of fuel released will simultaneously lubricate the surfaces that slide against each other.

One embodiment of the invention will now be explained in further detail with reference to the drawings in which:

- figure 1 shows a partial, longitudinal section through the fuel injector according to the invention,

- picture 2 on a larger scale, they show the part from picture 1 and there is drawn the compression spring with the related parts, and

- Figure 3 shows a diagram of the relationship between the engine load and the nozzle opening pressures of the known type and those according to the invention.

Figure 1 shows a fuel injector 1, generally constructed, having an outer housing 2 for mounting in a cylinder head. The housing is extended and at its upper end has a mounting part 3 which protrudes from the side and is fixed in the cover by means of screws and presses the contact surfaces 4 at the lower end of the housing against the corresponding contact surface formed in the cover, a fuel pump, not shown, or a similar source which periodically supplies highly compressed fuel, it is connected through a pressure pipe to the fuel insert 5 at the top of the injector, from where the fuel passage 6 goes centrally through the injector down to the nozzle 7 with a central cavity 8, from which the nozzle holes, not shown, expand radially for fuel injection into the engine cylinder.

The fuel passage may have a valve that opens to circulate preheated fuel to the injector between injection periods. The fuel passage goes through a stationary part in the form of an inserted part 9 which is in upward contact with a part 10 which is stationary in the housing, and is in downward contact with an intermediate part 11 which is pressed firmly against the slide guide 12 in the housing. The valve slide 13 is received so as to be longitudinally movable in the central guide bore 12 in the slide guide, and by its one end limits the downwardly projecting cylindrical part 11'' to the intermediate part. The guide bore centers the slide so that the annular conical abutment surface 14 located at the lower end of the valve slide and shaped like a needle, is coaxial with the corresponding valve abutment on the slide guide 12. When the valve slide is in its closed position with the abutment surface pressed against the valve abutment on the guide slide, the tip of the needle protrudes into the central cavity of the nozzle and here it is exposed to the pressure of the engine cylinder, which spreads through the holes of the nozzle into the cavity and acts on the valve slide with a force in the opening direction.

The lower end of the valve slide and guide, the slide 12 delimits the pressure chamber 15 which is connected to the fuel passage 6 through the inclined bores 16. Downward, the ring and the surface of the valve slide, which is inwardly limited by the needle, is exposed to the influence of the fuel pressure in the chamber 15 , because the fuel acts on the valve slide with a force in the direction of opening. The opening area of the valve slide is mainly determined by the difference in diameter between the outer diameter of the cylindrical part 11' and the inner diameter of the guide hole 12'.

The valve slide is also located in the closing direction, i.e. in the downward direction towards the valve seat, by means of the compression spring 17, the upper end of which is in contact with the upper spring guide 18 mounted movably on the inserted part 9, and the lower end of which is supported over the lower the spring guide 19 is also movably guided on the inserted part 9, by means of a grooved insert liner, the lower part of whose surface is in contact with the upper ring on the valve slide 13. The spring force is thus transmitted via the spring guide 19 and the inserted groove 20 to the valve slide 13. Annular the first piston 21 is mounted axially movable around the upper section of the insert part 9. The inner diameter of the sliding surface 22 (Fig. 2) of the piston is adapted to the opposite surface of the guide 21 on the insert part in such a way that the annular gap between the surfaces is close enough to the piston to sealingly surround the insert part .

The surface of the guide 23 ends at the bottom in a cylindrical recess made in the insert part and through the channel 24 is in contact with the central fuel passage 6 in the insert part. The recess merges down into a cylindrical part 25 of a smaller outer diameter than the guide surface 23. Below the part 25, the inserted part has an annular conical lower abutment part 26.

The first piston 21 has a lower inner ring 27 which has a conical abutment part 26' which can pressurize and contact the abutment part 26. In the area parallel to the recess and the cylindrical part 25, the first piston and the inserted part limit the first pressure. chamber 28 with an effective cross-sectional area determined by the diameter difference between the section 25 and the surface of the guide 23. The effective cross-sectional area is located on the upper surface of the ring 27, i.e. on the first surface facing away from the spring so that the fuel pressure supplied through the channel 24 to the first pressure chamber acts on the first piston is forced down.

The second piston 29 is formed integrally with the upper spring guide 18 and includes an annular intermediate part 30 that supports the lower tubular wall 31 and the upper tubular wall 32. The inner wall of the lower wall 31 is pressure-sealing and longitudinally movable in contact with the cylindrical second surface of the guide 33 of the inserted part 9, and the inner side of the upper wall 32 is in pressure-sealing and axially movable contact with the outer wall of the first piston 21. The first and second pistons together with the inserted part 9 delimit the second pressure chamber 34 with an effective cross-sectional area determined by the difference in diameter between the cylindrical inner sides of the lower wall 31 and the upper walls 32. A cylindrical recess formed in the outer side of the insert and located immediately below the adjacent part 26 forms the inlet connection 35 between the two pressure chambers, when the first piston moves away from the adjacent part 26.

The lower side of the ring 27 may have one or more protrusions or an annular protrusion with notches that hold a part of the second pressure chamber 34 located radially aligned with the protrusion in the inlet connection with the inlet connection 35, when the protrusion contacts the upper side of the intermediate part 30. In an alternative embodiment, which is not shown, the protrusion can be annular, and the inner side of the lower wall 31 can have a smaller diameter than the inner side of the ring 27 so that the part of the second pressure chamber located closest to the inlet connection 35 has an upward effective cross-sectional area that is open to the inlet connection 35 when the protrusion on the ring 27 touches the upper side of the intermediate part 30 and blocks the connection to the remaining part of the pressure chamber 34. With a suitable increase in pressure in the inlet connection 35, this effective surface will cause the second piston 29 to move away from the first piston, simultaneously revealing the complete effective cross-sectional area of the second pressure chambers.

The second pressure chamber 34 is constantly in contact with a limited discharge passage consisting of a pressure-sealing annular slot between the inner side of the upper wall 32 and the cylindrical outer side of the first piston and a pressure-sealing annular slot between the inner side of the lower wall 31 and the guide surface 33 on the inserted part .

Now follows a description of how the two pistons automatically create the desired spring force in the compression spring 17. When the engine is stopped and the fuel passage 6 is not under pressure, the two pistons are in the position shown in the drawing where the compression spring 17 with its factory preload presses the other piston 29 upwards to touch the first piston 21, which transmits the spring force to the inserted part 9 through the adjacent parts 26 and 26

When the engine is started and the load increases, the pressure in the fuel passage 9 increases during each injection period to a maximum pressure which at low loads corresponds to the injector opening pressure and at higher loads is determined by the flow resistance to the nozzle holes. The maximum pressure in the fuel passage thus increases with increasing engine load.

The pressure in the fuel passage 6 spreads through the channel 24 towards the first pressure chamber 28, and when the pressure here reaches a level at which the downward force on the first piston exceeds the existing spring force, the first piston is driven against the spring, which is compressed between the spring guide 18 and 19, and at the same time the fuel flows through the flow connection 35 into the second pressure chamber where pressure is created to a level that brings the first piston 21 back into contact with the adjacent part 26, while the second piston 29 remains in the position in which the spring has given a special load.

If the pressure in the fuel channel 6 rises to a higher level during the following injection periods, the movements of the piston are repeated so that the spring 17 gives a load that is linearly dependent on the maximum pressure in the fuel passage 6.

Through the pressure-sealing annular slits, a small amount of fuel is constantly drained from the second pressure chamber, the fuel goes to the discharge hole, which is not shown, through the inner cavity into the housing 2.

In this way, the fuel injector according to the invention achieves a spring force and therefore an opening pressure that increases with increasing engine load, as seen in Figure 3. This enables a lower opening pressure at low engine loads, because the injector automatically creates a high; opening pressure required at full load. Therefore, the compression spring can be pre-made with a preload that contributes to the opening pressure at low loads of approx. 200 bar, which contributes to steady operation of the engine at partial loads, and at the same time, the opening pressure at full loads is higher than with known injectors, which contributes to good fuel atomization at the beginning of the injection period.

Instead of an upper central direction, the fuel passage 6 may in a well-known manner include a number of channels which extend into the stationary intermediate part along the outside of the spring and which open into the pressure chamber 15 via oblique channels in the slide guides. With this design, the opening area of the valve slide is defined by the lower annular end surface surrounding the needle.

Claims (7)

1. A fuel injector (1) for an internal combustion engine, particularly for a large two-stroke diesel engine, having an extended outer housing (2) for mounting in the cylinder head and a fuel passage (6) opening into the nozzle (7), and has a valve slide (13) which can be moved longitudinally in the slide guide (12), which slide acts up in the direction towards its abutment of the valve by means of a preloaded compression spring (17) and in the opposite direction by means of fuel pressure in the fuel passage, the first piston ( 21), which can be moved axially in the housing, located at the end of the pressure spring and has a first surface facing away from the spring and, together with the fixed part (9), defines the first pressure chamber (28) which is connected to the fuel passage by of the channel (24), said pressure spring acts on the first piston in the direction towards the end position with the minimum volume of fuel in the first pressure chamber, characterized by the fact that the spring is connected by a force-transmitting connection to the second piston (29) which has a second surface facing away from the spring and forms the end of the wall in the second pressure chamber (34), so that when moving in the direction from the said final position, the first piston (21) opens the inflow connection (35) between the first and second pressure chambers, so that the second pressure chamber (34) has a larger effective cross-sectional area but the first chamber (28), and that a limited discharge passage connects the second pressure chamber with the discharge. 2. A fuel injector according to claim 1, characterized in that the first piston (21) has an abutting part (26') that blocks the inflow connection (35) between the first and second pressure chambers by contacting the corresponding abutting part (26) on the stationary part. 3. Fuel injector according to claim 1 or 2, characterized in that the discharge passage is of such a limited size that the discharged volume at full engine load during the engine cycle comprises from one half to one twentieth of the fuel volume in the second pressure chamber (34). 4. A fuel injector according to any one of claims 1-3, characterized in that the effective cross-sectional area of the first pressure chamber (28) is smaller than the opening area of the valve slide. 5. A fuel injector according to any one of claims 1-4, characterized in that the second pressure chamber (34) has an effective cross-sectional area several times larger than that of the first chamber (28). 6. Fuel injector according to any one of claims 1-5, characterized in that, in a known manner, the compression spring (17) is mounted between two spring guides (18, 19) which can be moved longitudinally on the centrally inserted part ( 9) which is stationary in the housing, that the second piston (29) is formed in the upper guide of the spring (18) which is located opposite the nozzle (7) and has a lower tubular wall (31) which surrounds the inserted part by pressure sealing, and the upper tubular wall (32) which has a larger inner diameter than the lower wall and pressure sealingly surrounds the first piston (21), and the intermediate part (30) which connects the walls, whose upper surface forms the second surface, that the first piston (21) is tubular and is enclosed between of the inserted part and the upper wall of the second piston and has a lower ring (27) whose upper surface forms a first surface, which merges inwardly into the adjacent part (26'), and that the corresponding adjacent part (26) of the inserted part faces downwards and is located between the canal that passes skr oz (24) into the fuel passage (6) and the lower surface of the reduced diameter forming the inlet junction (35) between the two chambers. 7. Fuel injector according to claim 6, characterized in that the discharge passage consists of pressure-sealing annular slits between the two walls (31, 32) of the second piston, i.e. between the first piston (21) and the inserted part (9).