It is previously known to use one or more cylinders in diesel engines as compressor cylinders by providing each of the cylinders with an extra exhaust valve and equipping the engine injection system with devices through which the fuel supply can be cut off to the engine cylinders functioning as compressor cylinders. Such a compressor arrangement can i. a. be designed to supplement the ordinary engine unit air compres- sor, which normally supplies the vehicle compressed air system when its capacity is not sufficient for some reason, or it can function as the sole compressed air source for the vehicle.

SE 9003735-9 describes an internal combustion engine of the type described by way of introduction. The extra exhaust valve is hydraulically operated with the air of a hydraulic pump with control valves. When the engine cylinder with the extra exhaust valve is to work as a compressor, the valve is set in the open position and the fuel supply via the associated unit injector is cut off by holding its spill valve open. The exhaust valve is kept constantly open as long as the compressor function is to be maintained, and re-entry of air via the extra exhaust valve is prevented during the intake stroke by means of a non-return valve disposed in the exhaust channel.

An arrangement of this type functions satisfactorily when the engine is to work as a compressor at low load, for example, during engine braking when driving downhill. If an extra air supplement is required even when the engine is loaded, several problems

will arise, however. The vibrations arising when an engine cylinder is completely decoupled, can be so great that the vehicle gearbox in particular will be subjected to unacceptable wear, which reduces its useful life to less than half of the normal useful life. If the engine is turbo-charged, there is lost not only a sixth of the normal power (in a six-cylinder engine with one cylinder decoupled) but as much as half of the power can be lost due to the loss of charge air pressure which occurs when the gas flow to the turbine of the turbocompressor is reduced. An additional problem is that the high air pressure in the cylinder in combination with the constriction of the non- return valve generates very high temperatures in the non-return valve, placing heavy requirements on its design to make its function and useful life acceptable.

The purpose of the present invention is in general to provide an internal combustion engine of the type described by way of introduction, in which the compressor function can be utilized even when the engine is loaded, without problems arising with vibra- tions and heavy power losses.

This is achieved according to the invention by virtue of the fact that the plunger of the plunger cylinder device is subjected to the pressure in a pump chamber which commu- nicates with the fuel injection system and which belongs to pump means which are driven by a cam element having a cam curve, which initiates a pump stroke in said pump means before cam means driving pump means in the unit injector initiate a pump stroke, and that the valve means comprise a control valve which, in its closed position, blocks the communication of the pump chamber of the first mentioned pump means with the fuel injection system, so that the fuel volume enclosed during the pump stroke produces a displacement of the plunger of the plunger cylinder device with asso- ciated displacement of the disc valve towards it open position.

By using the fuel, instead of the engine oil, as a hydraulic medium and controlling the fuel in the same manner as the injection system, a very rapid engagement and dis- engagement of the compressor function can be achieved. Thus, the control valve can be an electrically operated spill valve of the same type as the unit injector spill valve

and can also be controlled by the same control unit that controls the fuel injection.

The rapid engagement and disengagement of the compressor function, i. e. the opening and closing of the extra exhaust valve, can be determined however desired by the software in the engine-controlled computer in the same manner as the fuel injection is determined. This rapidity means that the compressor function does not need to be engaged during each cycle during power extraction from the engine; only at certain intervals, e. g. each fifth cycle. This in turn reduces engine vibrations and allows the charge pressure to be kept high, thus also minimizing power losses. If, for example, one engine cylinder in a six-cylinder engine works as a compressor each fifth cycle, the loss will be 20% of slightly more than 16%, or ca 3,5%. This loss can be easily compensated for with a somewhat increased amount of fuel. Finally, the rapid control (within milliseconds) of the extra exhaust valve provides complete force control/time control of the valve, so that the non-return valve and the problems associated therewith can be eliminated.

The invention will be described in more detail below with reference to examples shown in the accompanying drawings, where Figure 1 shows schematically a cross- section through a first embodiment of a portion of an engine according to the inven- tion, Figure 2 shows a corresponding cross-section through a second embodiment of a portion of an engine according to the invention, and Figure 3 shows a cam profile of a cam element cooperating with the unit injector.

1 in Fig. 1 designates a portion of a cylinder block, and 2 indicates a cylinder head.

Parts such as intake and exhaust ducts with associated valves and other components which are not directly related to the invention, have been eliminated for the sake of simplicity. The engine is a direct injection diesel and has a piston 3 in a cylinder 4.

It has an injection system with a so-called unit injector 5 of a type known per se and comprises a pump portion 6 with a pump plunger 8 in a pump chamber 7 and an injector portion 9 with a valve needle 10 controlled by the fuel pressure. The pump chamber 7 communicates via a spill valve 11 with a fuel channel 12 on the low- pressure side of the fuel system, to which fuel is fed by a feeder pump (not shown).

the spill valve 11 has a valve element 13 operated by an electromagnet and it is con- trolled in a manner known per se by the engine control computer (not shown). With <BR> <BR> <BR> the aid of the spill valve 11 the injection time and fuel amount are determined. The pump stroke of the injection pump plunger 8 is achieved with the aid of a rocker arm 14 and a cam element 15, cooperating therewith, on a cam shaft 16. A return spring 17 returns the pump plunger 8 to a starting position when the cam follower 18 on the rocker arm 14 reaches the base circle of the cam element (see Fig. 3). In the position shown in Fig. 1, the spill valve 11 is closed, which means that a channel 19 from the pump chamber 7 is shut off from a channel 20 to the low-pressure side of the fuel system. When a certain pressure (ca 300 bar) has been reached during the stroke of the pump 8, the fuel needle 10 opens the fuel nozzle and the injection begins and is maintained until the spill valve 11 opens, connecting the channel 19 to the channel 20 to the low-pressure side 12, and the injection will be stopped even if the pump stroke has not been completed.

The cylinder shown has an extra exhaust valve 21 in addition to the exhaust valve which is not shown. The valve 21 has a spindle 22 in a bore 23, and a valve disc 24 which rests against a seat 25 in a broader portion 26 of the bore 23. From the broader portion 26 a channel 27 exits the cylinder head into a conduit 28a which opens into a pressure tank 28b.

The valve disc 24 is pressed against its seat 25 by a helical spring 29, which is tensioned between a shoulder 30 in the cylinder head 2 and a plate 31 solidly joined to the spindle 22.

A pump device 32 corresponding to the pump portion 6 of the injector 5 is arrange in the cylinder head 2 to one side of the injector 5. The pump device 32 has a pump plunger 34 disposed in a pump chamber 33. The piston movement is achieved with the same rocker arm 14 and cam element 15 as operate the plunger 8 in the injector 5. An operating plunger 35 in a cylinder chamber 36 has an end 37 abutting against the end <BR> <BR> <BR> of the valve spindle 22 and its opposite end 38 protruding into the pump chamber 33,

which can be connected via a channel 39, a control valve 40 and channels 41 and 42, to the low-pressure side 12 of the fuel system. The valve 40 is of the same type as the spill valve 11 and thus has a valve element 43 which is electromagnetically operated, preferably by the control computer (not shown) of the engine, as is the spill valve 11.

In the position shown in Fig. 1 of the valve element 43, the valve is closed and the pump chamber 33 is thus cut off from the low-pressure side 12 of the fuel system, where the pressure is normally about 4 bar. The force of the return spring 29 is thus adapted so that it keeps the exhaust valve 21 closed when there is this pressure in the pump chamber. When the valve 40 is closed, there will be an increasing pressure in the pump chamber 33 during the stroke of the pump plunger 34, thus creating a force against the operating plunger 35, which overcomes the force of the return spring 29 so that the exhaust valve 21 opens.

The pump movements of both the unit injector 5 and the pump device 32 are achieved with the aid of the cam element 15, the cam profile of which is shown most clearly in Fig. 3. By merely grinding down the basic circle"a"of an"ordinary"cam in a fuel injection system with unit injectors, a smaller base circle"b"is obtained, which makes it possible to form an extra cam curve"c"for the stroke of the pump plunger 34 of the exhaust valve 21, in front of the ordinary cam curve"d"for the stroke of the pump plunger 8 of the unit injector 5. The cam curve and are so arranged in relation to each other that the extra exhaust valve is opened after the initial portion of the piston compression stroke. In Fig. 3, a designates the cam angle for the stroke of the exhaust valve 21 and (3 designates the cam angle for the stroke of the injector 5. During engine operation, the plungers 8 and 34, respectively, of the unit injector and the exhaust valve, respectively, will perform one stroke per work cycle. The opening and closing of the exhaust valve 21 is only regulated with the aid of the valve 40, which is control- led by the engine control unit. Thus, the exhaust valve 21 can be controlled as rapidly as the fuel injection, i. e. within milliseconds. In a practical embodiment, the stroke for opening the exhaust valve can be ca 4 mm, which is to be compared with the fuel injection pump stroke of ca 17 mm. The rapid control of the exhaust valve makes it

possible, depending on the operating conditions (light load or heavy load), with the aid of the engine control unit, to open and close the exhaust valve during each work cycle or at a chosen interval, e. g. each fifth work cycle. It is possible, at least theoretically, to only use an initial portion of the compression stroke for air pumping and then ignite the remaining air after the final portion of the compression stroke and thus reduce the engine power loss further if this should be required.

In the embodiment described above, the control of the exhaust valve 21 is, in principle, independent of the setting of the unit injector spill valve 11 even though it is, of course, in practice so that if the entire compression stroke is to be used for air pumping, the spill valve 11 must be open to prevent fuel injection.

Fig. 2 shows a second embodiment of an engine according to the invention. Parts with direct counterparts in Fig. 1 have been given the same reference numerals as in Fig. 1, and only the structural and functional differences between the two embodiments will be described here.

In the embodiment in Fig. 2, there is no separate pump arrangement 32 for the exhaust valve 21. Instead, the operating plunger 35 of the exhaust valve 21 extends into a chamber 50 which communicates via a channel 51 with the pump chamber 7 of the unit injector pump portion 6, which is used here both to pump fuel and to open the exhaust valve 21. The opening and closing function is the same as described above.

The difference lies in the structural simplification, i. e. the elimination of an extra pump and the integration of the two pump functions in the fuel pump 6. This requires the spill valve of the unit injector 5 to be involved in the control of the exhaust valve 21, since the spill valve also must be closed for high pressure to initially be generated in the pump chamber 7 and thus in the chamber 50 as well. The plunger areas are selected so that when the pressure at the beginning of the pump stroke reaches ca 150 bar (i. e. about half of the injector opening pressure of ca 300 bar), the exhaust valve 21 will open. The pump plunger 8 also has a valve function here by virtue of the fact that when it covers the outlet through the channel 51, it cuts off the connection between the

pump chamber 7 and the chamber 50 of the operating plunger 35. The spill valve 11 can now be opened to prevent fuel injection without thereby reducing the pressure in the chamber 50. The subsequent closing of the exhaust valve after a completed com- pression stroke is effected by merely opening the control valve 40 and connecting the chamber 50 to the low pressure side of the fuel system.

The invention provides, with simple and mostly already existing means, a previously unachieved speed and precision in opening and closing an extra exhaust valve during compressor operation of a diesel engine.