AT402432B - INTERNAL COMBUSTION ENGINE - Google Patents

Description

AT 402 432 B

The invention relates to an internal combustion engine in which an intermediate store is provided for each cylinder, into which gas is transferred from the cylinder after the end of the combustion in a phase of high pressure in the cylinder via a connection channel controlled by an overflow valve to the cylinder, and the intermediate container in a phase of low pressure in the cylinder, the gas is fed back to the cylinder at the beginning of the compression, the connection channel between the cylinder and the intermediate store opening above the piston in its top dead center position directly into the combustion chamber and the flow connection between the intermediate store and the cylinder being able to be established independently of the exhaust valve control.

In known internal combustion engines, gas is withdrawn from the cylinder in a phase of high pressure and transferred to an intermediate container. This gas is used either to introduce the fuel into the combustion chamber (mixture injection) or, in the case of direct fuel injection into the combustion chamber, to improve the combustion by blowing the gas into the injection jet. In both cases, the storage volume is relatively small compared to the stroke volume with regard to the task that it has to perform.

From DE 35 33 014 C2 an internal combustion engine of the type mentioned is known, in which the stored air is returned to the combustion chamber as a function of the injection pressure. This is to achieve the most complete possible combustion of the fuel and thus to reduce harmful emissions. The gas is withdrawn from the cylinder into the intermediate store and returned from the intermediate store via the injection nozzle into the cylinder via check valves, which means that the removal and return is strongly dependent on the piston position and the injection pressure. Removal and return independent of the piston position and the injection pressure is not possible.

From GB-A 3 790 a.d. A submarine internal combustion engine is known in 1906, in which an intermediate store is connected to the cylinder via a connection channel controlled by an overflow valve. Here, too, gas is transferred from the cylinder after the end of the combustion in a phase of high pressure in the cylinder, stored temporarily in the intermediate container and fed back to the cylinder via the connecting channel in a phase of low pressure. However, the connecting channel opens into the outlet channel, so that the flow connection between the intermediate store and the cylinder is only possible when the outlet valve is open.

DE 228 997 C describes an internal combustion engine in which the combustion air is heated directly by combustion products from an earlier stroke. This is done in such a way that at the end of the working stroke, combustion products are transferred to a sensor, from which they flow into the cylinder at the end of the suction stroke or at the beginning of the compression stroke. The channel leading to the transducer is controlled by the piston. This results in very short control times and, due to the rigid relationship between piston movement relative to the channel, the pressure level in the transducer is limited, which means that a large storage volume is necessary. The same applies to a mixture-compressing four-stroke internal combustion engine known from DE 521 920 C, in which openings in the cylinder leading to a container are controlled by the piston.

In an internal combustion engine designed according to DE 1 751 473 A, the generation of a pressure gradient by throttling in the exhaust pipe or intake manifold results in only a low pressure level, similar to known exhaust gas recirculation systems. A major charging effect cannot be achieved in this way.

From US Pat. No. 3,799,130 A, an internal combustion engine is also known with a connecting channel starting from the cylinder and opening into an intermediate store, the opening of the connecting channel into the piston being controlled by the piston position. US 4,282,845 A also uses this principle. In addition, the entry into the buffer store can be controlled via its own lift valve. Nevertheless, there is a direct dependency on the piston position.

All these internal combustion engines have the disadvantage that the removal and supply of the gas from the store or into the store depends on the position of the piston of the internal combustion engine and control of the removal and supply of the gas is therefore not possible without restriction. Due to the removal and return that is rigidly dependent on the piston position, optimal conditions can only be achieved in individual operating areas of the internal combustion engine. In other areas of operation, the rigid design adversely affects efficiency, fuel consumption and pollutant emissions.

The present invention has for its object to avoid these disadvantages and to improve the efficiency of the internal combustion engine, its fuel consumption and its pollutant emissions in an internal combustion engine of the type mentioned. According to the invention, this is done in that the connecting channel between the intermediate store and the cylinder is the only flow connection to the intermediate store and the overflow valve in a manner known per se independently of the piston position and depending on at least one operating parameter of the internal combustion engine from group 2

AT 402 432 B

Load, speed, combustion chamber pressure and combustion chamber temperature is controllable. The advantages of this intermediate gas storage and recirculation at the beginning of the compression stroke are the achievement of a variable total charge mass in the combustion chamber with the fresh air charge mass being retained. A variable effective compression ratio can thus be achieved. For Otto engines, this means e.g. in partial load areas close to idle, an improvement in thermal efficiency of about 30%. In addition, the ignition delay in auto-ignition internal combustion engines is shortened by increasing the charge temperature. The nitrogen oxide emission due to an increased proportion of inert gas in the total charge is reduced. Another advantage is that the geometric compression ratio for the engine can be reduced without internal exhaust gas recirculation. The starting behavior is improved due to the high charge mass and charge temperature that are possible due to the residual gas thrust.

In the case of self-igniting, mixture-controlled internal combustion engines, the internal exhaust gas recirculation results in a partial charging effect in part-load areas. This is because, due to the high air numbers in the residual gas, there is still a sufficient amount of unburned oxygen.

The exhaust valve of the internal combustion engine can only be part of it, e.g. 90% of the possible total stroke opened and then very quickly up to a stroke of e.g. 10% are closed, with compressed gas in the cylinder of the internal combustion engine being pushed into the intermediate container via a control valve actuated by the camshaft, and the return to the cylinder before the commencement of the compression takes place via the control valve. This solution has the advantage that no phase shift to the crankshaft position due to different combustion chamber pressures is required for the control valve.

In a further embodiment of the invention it can be provided that the volume of the intermediate container is less than 50%, preferably 15 to 30% of the stroke volume of the cylinder. The advantages mentioned can be optimized by the relative size of the intermediate container.

The operating parameter used to control the overflow valve can preferably be the load and the speed of the internal combustion engine or the combustion chamber pressure and the combustion chamber temperature of the internal combustion engine.

When the overflow valve is controlled as a function of the combustion chamber pressure of the internal combustion engine, according to a further feature of the invention, the actuation of the overflow valve can be phase-shifted to the course of the pressure in the cylinder.

The invention is explained in more detail below with reference to the drawings. 1 shows a schematic illustration of an internal combustion engine for carrying out the method according to the invention in section through the cylinder axis, FIG. 2 shows a function diagram for the internal combustion engine according to FIG. 1, FIG. 3 shows a detail of an internal combustion engine according to the invention and FIG. 4 shows a spring diagram. 5 shows functional characteristics of a further embodiment of the internal combustion engine according to the invention.

In the four-stroke internal combustion engine shown in FIG. 1 according to the invention, the cylinder is denoted by 1, the piston by 2 and the cylinder head by 3. The charge change of the combustion chamber 4 takes place via the inlet channel 6 controlled by the inlet valve 5 or via the outlet channel 7, which is controlled by the outlet valve 8. An intermediate store 9 is arranged in or in connection with the cylinder head 3 and is connected to the combustion chamber 4 via the connecting duct 10. The overflow valve arranged in this connecting channel 10 is designated 11. The volume of the intermediate store 9 is less than 50% of the stroke volume of the cylinder 1, but preferably 15 to 30% thereof.

The overflow valve 11 is controlled in such a way that the gas is removed from the cylinder after the end of the combustion, fed to the intermediate container 9 and fed back to the combustion chamber 4 from the intermediate container 9 before the start of compression. Arrows 12 and 13 symbolize the inflow and outflow of gas into the intermediate container 9 or from this intermediate container.

Fig. 2 shows the internal combustion engine according to Fig. 1, plotted over the crank angle for two crankshaft revolutions, the valve lift 8 'of the exhaust valve 8, the valve lift 5' of the intake valve 5 and the valve lift 11 'of the overflow valve 11 in mm. In addition, the cylinder pressure 14 is plotted over the crank angle. In the first crank angle section 0 to 180 ·, which is denoted by 15, the combustion or expansion takes place in cylinder 1. In the second section 16, which extends to the top dead center, the gas is pushed out, in the third section 17 the suction takes place and in the fourth section 18 to the top dead center, the compression takes place.

As can be seen, the valve lifts of the exhaust valve and the intake valve are approximately the same size and overlap only slightly at 360 * crank angle. In the crank angle range 19 between approximately 60 * and 100 *, the overflow valve 11 is raised approximately 2 mm by a cam seated on a camshaft rotating at half the engine speed, a connection being established between the combustion chamber 4 and the intermediate store 9. The same happens in the crank angle range 20 between 580 * and 620 * by a second cam. In the first opening period, before the end of section 15, where a third

AT 402 432 B If there is overpressure in the cylinder, gas flows from cylinder 1 into the intermediate store 9. Conversely, in section 18, in the crank angle region 20, the compressed gas flows from the intermediate store 9 into the combustion chamber 4, so that an additional charge compression takes place while increasing the cylinder content.

The overflow valve 11 can be operated in various ways (pneumatic, mechanical, electrical, hydraulic, etc.).

With constant control times for the overflow valve 11, the pressure in the intermediate store 9 is a function of the engine operating state. Depending on the load and speed, the cylinder pressure curve in the expansion phase is different.

In order to take this phenomenon into account, if necessary, the overflow valve 11 can be actuated in a phase-shifting manner in relation to the cylinder pressure curve. However, the control mechanism for this overflow valve 11 can also be made dependent on the load, the speed or on pressure and temperature in the cylinder.

A simple mechanical embodiment of a control of the overflow valve 11 which is dependent on the pressure in the cylinder is shown in FIG. 3 and its function is explained in more detail in FIG. 4.

The overflow valve 11 shown in FIG. 3 controls the connecting channel 10 and is designed as a poppet valve. The effective area of the valve plate 21 is designated A. The overflow valve has on its end opposite the valve plate a hollow cylindrical extension 22 which is fixedly connected to the stem 24 of the overflow valve 11 via its intermediate base 23. A helical compression spring F2, which loads the overflow valve 11 in the closing direction, engages on the underside of the intermediate base 23. Above the overflow valve 11 is a camshaft 25, the axis 25 'of which is the axis 11 " of the overflow valve 11 cuts. This camshaft 25 rotates at half the engine speed and has two cams 26 and 27 offset by approximately 180 *, one of which is the overflow of gas into the intermediate store 9 in the expansion phase and the other is the blowing of the stored gas into the cylinder in the compression phase controls. The cams 26 and 27 work together with a tappet 28, which is also hollow-cylindrical and is pressed against the camshaft 25 by the helical compression spring F1, which is supported on the top of the intermediate floor 23, and is thus connected to the camshaft in a force-locking manner. The maximum stroke of the plunger 28 with respect to the hollow cylindrical extension 22 is denoted by H1, the maximum stroke of the hollow cylindrical extension 22 with respect to the contact surface 29 of the spring F2 is denoted by H2.

The total force from the valve force, determined by the valve disk surface A and the pressure p in the combustion chamber 4, and the spring force F2 is counteracted by the spring force F1, which is increased by the cams 26 and 27 from the basic position shown.

The force diagram is shown in FIG. 4. Here are the force curves in the direction of the axis 11 " of the overflow valve 11 effective forces F plotted over the stroke H of the overflow valve 11 or, respectively, the time t. A desired constant cylinder pressure value p at the beginning of the control valve stroke results when the spring force F1 is equal to the total force from the " cylinder force FZi, FZ2l ... " corresponding to the respective operating state of the internal combustion engine. and the spring force F2. The control valve 11 opens when

Fi = F2 + FZ! = f2 + fz2 etc. The desired constant opening force is denoted by C.

The mode of operation is, for example, as follows: from the closed position of the overflow valve 11 shown in FIG. 3, the tension of the spring F1 is increased when the camshaft rotates, so that the overflow valve 11 opens at a specific cylinder pressure p. When the spring force F1 decreases due to further rotation of the camshaft 25, the " cylinder force " and the overflow valve 11 is closed again. At higher loads it can easily be achieved that this overflow valve 11 does not open because the pressure p does not reach the low value required for this. The opening and closing of the overflow valve 11 in the points determined by the two cams 26 and 27 is therefore dependent on the pressure p in the combustion chamber 4 and thus on the operating state of the internal combustion engine.

In principle, the following stroke profiles are conceivable for both opening phases of the overflow valve 11 controlled by the cams 26 and 27: 4

Claims (5)

AT 402 432 B 1.) With a cam stroke of less than H1, the opening of the overflow valve is only determined by the cylinder pressure p + spring F1 and can be omitted for certain operating conditions - for example in full load operation. 2.) In the event of a cam stroke between H1 and H1 + H2, the overflow valve 11 is opened at a certain point in time specified by the cam contour. However, this forced opening - depending on the prevailing cylinder pressure p - is preceded by an opening controlled by the cylinder pressure and / or follow. 3.) Finally, it is also conceivable to form asymmetrical cams 26 and 27, which differ both in their contour and in their cam stroke in order to compensate for different pressure conditions during the overflow of the gases into the intermediate store and their blowing into the cylinder through different opening times can. The method shown in Fig. 5 differs from that according to Fig. 2 in that the outlet valve 8 is only e.g. 90% of the total stroke opened and then very quickly up to a stroke of e.g. 10% is closed. Due to the pushing-out effect of the piston 2 of the internal combustion engine, a pressure p is built up in the cylinder 1. This is pushed into the buffer 9. In this case too, an overflow valve 11 is present, which, however, controls the inlet into the intermediate store much later, in comparison with the method according to FIG. 2, namely between 240 and 280 * crank angle. The crank angle range for the inflow into the intermediate store is designated by 29. From the beginning of the overflow at a crank angle of 240 *, the cylinder pressure curve 14 shows a slight increase because the part of the load which has not yet been pushed out is somewhat compressed again. The overflow to the intermediate container is in turn controlled by a camshaft, analogously to the embodiment according to FIG. 2. The overflow from the intermediate container 9 to the cylinder during the intake stroke must also be controlled by a cam, analogously to that shown in FIG. 2. The effect of such an internal exhaust gas recirculation was thermodynamically calculated for a diesel engine with a cylinder displacement of 0.53 l. This resulted in a partial load point of pmi = 3.48 bar at 1300 rpm. Geometric compression ratio standard 20.5: 1 standard with exhaust gas recirculation 20.5: 1 standard lowered 16: 1 buffer volume 0 0.13 0.13 charge mass (g) 0.521 0.576 0.584 effective compression ratio 20.5 24.7 20.5 charge temperature (* K) 1036 1226 1155 ignition delay (crank angle) 5.4 3.4 4.1 Es applies that with the reduced geometric compression ratio of 16: 1 in partial load with internal exhaust gas recirculation the same effective compression ratio of 20.5: 1 is used as for the standard starting point. The charge temperature rises to 1155 * K compared to 1036 * K, which corresponds to a shortening of the ignition delay by 1.3 * crank angle from 5.4 to 4.1 crank angle. The direction of nominal power can be driven without internal exhaust gas recirculation with the geometric compression ratio of 16: 1. 1. Internal combustion engine, in which an intermediate store is provided for each cylinder, in which gas is transferred from the cylinder to the cylinder via a connection channel controlled by an overflow valve in a phase of high pressure in the cylinder after the end of combustion and the intermediate container in a phase of low pressure Cylinder supplies the gas to the cylinder again at the beginning of the compression, the connection channel between the cylinder and the intermediate store opening above the piston in its top dead center position directly into the combustion chamber and the flow connection between the intermediate store and the cylinder being able to be established independently of the exhaust valve control, characterized in that the connecting channel (10) between the intermediate store (9) and the cylinder (1) is the only flow connection to the intermediate store (9) and the overflow valve (11) in a manner known per se independently of the piston position and depending on at least one ns an operating parameter of the internal combustion engine from the group load, speed, combustion chamber pressure and 5 AT 402 432 B combustion chamber temperature can be controlled. 2. Internal combustion engine according to claim 1, characterized in that the volume of the intermediate store (9) is less than 50%, preferably 15 to 30% of the stroke volume of the cylinder (1). 3. Internal combustion engine according to one of claims 1 or 2, characterized in that the operating parameters for controlling the overflow valve (11) are the load and the speed of the internal combustion engine. 4. Internal combustion engine according to one of claims 1 or 2, characterized in that the operating parameters for controlling the overflow valve (11) are the combustion chamber pressure and the combustion chamber temperature of the internal combustion engine. 5. Internal combustion engine according to claim 1, wherein the at least one operating parameter is the combustion chamber pressure, or according to claim 4, characterized in that the actuation of the overflow valve (11) to the course of the pressure in the cylinder (1) is phase shiftable. Including 2 sheets of drawings 6