DE4333424B4 - Method for controlling a diesel engine - Google Patents

Abstract

Method for controlling a diesel engine with the following steps: detection of an engine operating state,
Controlling a fuel injection in accordance with the detected operating state of the engine,
Recirculating an exhaust gas via an exhaust gas recirculation valve between an inlet duct and an exhaust duct,
Detecting an exhaust gas recirculation ratio (EGR ratio),
characterized in that the exhaust gas recirculation ratio (EGR ratio) is controlled as a function of the detected engine operating state and at the same time a fuel injection time (IT) of a fuel injector (20) is set as a function of the detected exhaust gas recirculation ratio (EGR ratio).

Description

The present invention relates to a method for controlling a diesel engine, in particular for controlling the combustion of the diesel engine, according to the preamble of the independent claim 1.

Out DE 34 28 371 A1 A method for measuring and controlling operating data of internal combustion engines is known, in which the exhaust gas composition, in particular the soot content of the exhaust gas, is detected and the exhaust gas recirculation and the start of spraying are carried out as a function of the detected soot content of the exhaust gas in order to improve the exhaust gas quality achieve. In this case, the exhaust gas recirculation ratio is only regulated as soon as a soot concentration exceeding a certain limit value is detected by the sensor device mentioned.

It has recently become known to provide diesel engines with an exhaust gas recirculation duct which is arranged between an intake air system and an exhaust gas system, an exhaust gas recirculation valve being arranged between an exhaust gas recirculation duct for setting a regulated exhaust gas recirculation rate in order to generate NO _x gas and suppress other harmful exhaust components, as described below.

From Japanese Patent Laid-Open No. Showa 60-162018, released on August 23, 1985, for example, is an intake air system for an internal combustion engine known with internal combustion, especially for a direct-injection diesel engine.

According to the above laid-open specification, the engine is provided with an exhaust gas recirculation system which serves to return an inert gas of the exhaust gas to the intake air system of the engine to prevent generation of NO _x , which is a harmful exhaust gas component of the exhaust gas. An exhaust gas recirculation valve is disposed in an exhaust gas recirculation duct (a duct for returning a portion of the exhaust gas into the intake air duct) of the exhaust gas recirculation system, and the exhaust gas recirculation valve is open in an engine operating condition area in which exhaust gas recirculation is required so that a constant amount of exhaust gas ( Exhaust gas recirculation gas) is mixed with the intake air, thus reducing a maximum temperature during the combustion of the fuel.

If the exhaust gas recirculation rate (_ (amount of exhaust gas recirculated / intake fresh air amount) x 100) [%]) becomes too large increases the smoke concentration in the exhaust gas. For this reason, according to the above Japanese Patent Laid-Open to enable a higher one Exhaust gas recirculation rate a vortex of the air / fuel mixture is reinforced accordingly.

If the vertebra is strengthened around a high exhaust gas recirculation rate to be able to realize the mixing of air and fuel is improved to a Reduce generation of smoke in the exhaust.

However, this eliminates the problem of smoke formation not solve satisfactorily. This means, it is difficult in the above-described diesel engine with increased swirl formation suppress the increase in smoke formation in the exhaust gas if the exhaust gas recirculation rate to an even higher one Value is set.

Specifically, if the exhaust gas recirculation rate is increased towards 100%, the concentration of NO _x in the exhaust gas is greatly reduced, but the concentration of smoke in the exhaust gas is greatly increased.

In this case, if a swirl ratio SR is increased, can significantly reduce the concentration of smoke in the exhaust gas become. Still exceeds the smoke concentration in an area in which the exhaust gas recirculation rate is high, a limit (for example, a legal requirement Limit). A reason for the reduction in smoke concentration in the presence of a increased Vortex is that the diffusion speeds of air and fuel during a so-called diffusion combustion is increased. As a result, can in a situation where the oxygen concentration is due to a high exhaust gas recirculation rate is small, this effect, i.e. the increased diffusion, no longer clearly occur, because there is a lack of oxygen in the air.

The swirl ratio SR is defined as follows: SR = V _C / N, where V _{C describes} the rotational speed of the eddy current in the tangent direction and N describes an engine speed (an engine speed).

In addition, when the swirl ratio SR is increased, the concentration of NO _x becomes larger.

The invention has for its object to improve a method for controlling a diesel engine of the type mentioned in such a way that the emission of NO _x and smoke in the exhaust gas of the engine is reduced in an operating condition range in which the combustion temperature due to an extremely prolonged ignition timing Delay period is reduced, the extended ignition timing delay period being triggered by a high exhaust gas recirculation rate in the exhaust gas recirculation system and thus a reduction in the oxygen concentration.

This object is achieved according to the invention in a method for controlling a diesel engine of the type mentioned at the outset by controlling the exhaust gas recirculation ratio as a function of the detected engine operating state and simultaneously a fuel injection timing of a fuel injector can be set depending on the detected exhaust gas recirculation ratio.

Further preferred configurations of the subject matter of the invention are set out in the subclaims.

The invention is illustrated below of embodiments and associated Drawings closer explained. In these show:

1 a schematic block diagram of a first embodiment of a device for controlling the combustion of a diesel engine;

2 a diagram of the exhaust gas recirculation rate based on the engine torque and the engine speed;

3 a cut one in 1 fuel injection pump shown;

4 a diagram similar to that in 1 shown diagram, however, the fuel injection timing is shown;

5 a flowchart for explaining the control of a fuel injection timing and a fuel injection interval according to the in 1 first embodiment shown;

6 a diagram in which a basic injection interval Avm is shown;

7 a diagram in which a fuel temperature correction quantity ΔItm _{1 is} shown;

8th a diagram in which a further fuel temperature correction quantity ΔItm _{2 is} shown;

9 a diagram in which both the concentration of smoke and NO _{x are shown in} relation to the exhaust gas recirculation rate;

10 a diagram of the concentration of NO _x and smoke based on the injection time;

11 a common diagram in which the cooling losses, the degree of constant volume and the fuel consumption of the diesel engine as a function of the regulated exhaust gas recirculation rate is shown;

12 a schematic block diagram of a second preferred embodiment of a device for controlling the combustion of a diesel engine;

13 a diagram of a gear ratio based on the engine speed;

14 a perspective view of a rotor blade position of the in 12 vortex control device shown when there is a strong vortex formation;

15 a perspective view of a rotor blade representation of the vortex control device when a small vortex formation occurs;

16 a diagram of a swirl ratio when the swirl control device is activated;

17 a flowchart for explaining a control of a swirl ratio by means of the swirl control device according to the in 12 shown second embodiment;

18 a diagram of the concentrations of HC, smoke, and NO _x based on the regulated exhaust gas recirculation rate in the case of the second embodiment;

19 a schematic block diagram of a third preferred embodiment of a device for controlling the combustion in a diesel engine;

20 a diagram of the oxygen concentration related to the engine operating conditions with respect to the intake air in the case of the third embodiment.

The following is a reference to the drawings taken for a better understanding for the invention to facilitate.

1 shows a first preferred embodiment of a method for controlling the combustion of a diesel engine.

How 1 shows, is the diesel engine designated at 21 with an intake air duct 23 , an exhaust duct 25 , an exhaust gas recirculation duct 26 which is between the exhaust duct 25 and the inlet air duct 23 is connected, and provided with a diaphragm exhaust gas recirculation valve which responds to a control vacuum.

A vacuum control valve 28 is used to set three levels of constant negative pressure, for which a negative pressure source is used, and the negative pressure is set on the basis of an actuation ratio signal, which is sent by the control unit 31 is issued.

For example, if there is a constant negative pressure directly into the exhaust gas recirculation valve 27 is introduced at a time when the maximum "out of service ratio" prevails (a ratio of out of service time at a constant time period), 50% of the exhaust gas is recirculated. This corresponds to an exhaust gas recirculation rate of 100%.

As the out-of-service ratio is gradually reduced, the exhaust gas recirculation valve is caused to decrease 27 supplied negative pressure that the exhaust gas recirculation valve opening angle is reduced, so that the flow rate through the exhaust gas recirculation channel 26 is reduced. In other words, when the out-of-duty ratio of the duty signal becomes smaller, the exhaust gas recirculation rate is reduced to 60% or further to 30%.

2 shows a diagram of the output torque (Nm) related to the engine speed (U / min) with the exhaust gas recirculation rate as a parameter.

In 2 the exhaust gas recirculation rate is 100% in a range of medium speed and medium engine load as well as in all engine load ranges with a low engine speed range.

Because in these operating areas, smoke generation is substantially zero at one Exhaust gas recirculation rate of 100% is depressed, deposition of smoke particles on the intake valve cannot occur for each cylinder. Deposition of smoke particles on the intake valve would occur if the smoke in the exhaust gas through the intake air duct 23 through the exhaust gas recirculation duct 26 would occur. In contrast, in an area with high engine speed and high engine load, such an increase in the combustion period occurs that the generation of smoke cannot be completely suppressed. In addition, because an increase in the exhaust gas temperature and an increase in the exhaust gas recirculation flow amount cause an increase in the intake temperature in this area, the effect of reducing NO _x due to the large exhaust gas recirculation rate in the exhaust gas recirculation system is reduced, which is why to avoid the exhaust gas recirculation rate gradually to 60% and Is reduced by 30%.

In order to regulate the exhaust gas recirculation rate in accordance with the engine operating conditions, the control unit has 31 a microcomputer installed in it. The control unit 31 regulates the amount of exhaust gas recirculation in stages based on a signal from a sensor 32 for detecting an opening angle of an accelerator such as a throttle opening angle (or depression angle of an accelerator pedal), a signal from an air flow meter 33 , a reference pulse (crankshaft angle) and a scale pulse (crankshaft angle), as will be described below.

To a characteristic of an exhaust gas recirculation rate (target exhaust gas recirculation rate), as described in 2 is shown, based on the engine torque and the engine speed, a table (not shown) is provided, which has the opening angle Acc of the engine throttle valve or the accelerator pedal (corresponding to the engine load) and an engine speed Ne as parameters, from which stored A value is taken from each table in order to derive a current target exhaust gas recirculation rate.

The exhaust gas recirculation rate flow amount at this time is calculated from the target exhaust gas recirculation rate and the intake air amount from the air flow meter 33 is recorded (fresh air flow quantity), this calculation being carried out as follows:
Exhaust gas recirculation flow rate = air flow rate according to the air flow meter x target exhaust gas recirculation rate ( 1 )

The out of service ratio of the operating ratio signal is finally determined to send the out of service signal to the negative pressure control valve 28 to transmit so that the exhaust gas recirculation gas into the intake air duct 23 flows with a predetermined exhaust gas recirculation flow amount.

In addition, there is a special design of the fuel injection pump 20 in 3 shown.

3 shows a distributor fuel injection pump 20 , the fuel injection quantity and the fuel injection timing being electronically controlled.

Such a fuel injection control system for a diesel engine is known, for example, from the publications DE 33 12 950 C2 or DE 32 38 697 A1 known.

How 3 shows is a drive shaft 4 with an output shaft of the engine 21 connected. An impeller feed pump 2 is by means of the drive shaft 4 driven. Fuel coming from the feed pump 2 drawn from a fuel inlet (not shown) becomes a pump chamber 5 inside a case 1 and to a piston chamber 12 a piston pump 3 by means of an intake duct 6 which communicates to the pump chamber 5 is open.

An engagement element 9a an end face cam 9 is firmly attached to a left end of the piston 7 mounted and is located with one end (as shown in 3 the right end) of the drive shaft 4 engaged so that a sliding movement in the axial direction of the piston 7 is possible. Both the end face cam 9 , as well as the piston 7 are on the same axial line as the drive shaft 4 arranged, the piston 7 is displaceable in the axial direction.

A roller cage is concentric on an outer periphery of a connecting part between the drive shaft 4 and the end face cam 9 arranged. The roller cage 10 is used to hold a plurality of roles 11 , In addition there are cam surfaces 9b formed, which serve as cam areas for different speeds, the number of which corresponds to the number of engine cylinders and wherein the cam surface 9b against the roles 11 by means of a spring 15 be pressed.

A plurality of suction slots 8th , the number of which is equal to the number of engine cylinders, are at a tip of the pistons 7 educated. If the cam surface 9b as in 3 is shown with the drive shaft 9 turns, it runs over the rollers 11 which in the roller cage 10 are arranged, and a reciprocating movement is carried out over a predetermined stroke, with fuel passing through the respective suction slot 8th was sucked in, and in the piston chamber 12 has been pressurized into the manifold ports for each of the cylinders which is connected to the piston chamber 12 are connected, and the fuel is thus directed to an injection nozzle by means of a supply valve.

A fuel return line 13 is between the piston chamber 12 and the pump chamber 5 connected, which is under a low pressure. An electromagnetic valve 14 which is very responds quickly is in the fuel return duct 13 provided, which valve is actuated in response to a signal (driving signal) which is supplied by a driving circuit in accordance with the operating conditions of the engine. The solenoid valve 14 is installed to control the amount of fuel, the injection timing and the injection duration. If the solenoid valve 14 during a piston pressure stroke 7 is closed, fuel injection is started and, conversely, when the solenoid valve is opened, the fuel injection operation is ended. In other words, the fuel injection start timing is controlled in accordance with a timing at which the solenoid valve 14 is closed, and the fuel injection amount is controlled according to a period of time during which the solenoid valve 14 closed is.

Although the NO _x concentration can be reduced as the exhaust gas recirculation rate increases, the smoke concentration is increased abruptly. In this case, even if the mixing is improved during the diffusion combustion due to an increase in the vortex, the concentration of the smoke can be insufficiently reduced at the large exhaust gas recirculation rate.

In order to eliminate the above problem, the control unit decelerates 31 the fuel injection start timing to a time or crank angle past a top dead center (TDC), so that a greatly increased ignition timing delay period occurs when the engine is operated in a range where a high EGR rate occurs.

4 shows a diagram of the fuel injection timing comparable to the diagram of 2 ,

How 4 shows, in an operating range in which there is a medium engine load (middle range) or in which the engine load is high while the engine speed range is low, the fuel injection point in time to a crankshaft angle behind the top dead center (+ 4 ATDC and + 2 ATDC) delayed. This is because the significant delay in fuel injection timing causes the intake air to be at a relatively low temperature, so that a ratio of premixed air-fuel combustion gas is increased, which results in smoke suppression.

When the engine operating conditions are in a medium to high engine load range, in which the engine speed is also medium or high, the fuel injection timing shifted to an earlier point in time with increasing engine speed.

This is because, though the ignition delay period is the crankshaft angle at which the ignition delay occurs (a value of the conversion from the ignition timing delay period to the crankshaft angle), larger in relation to an increase in engine speed (rpm), so that the injection timing with an increasing magnification of the Engine speed earlier takes place at the ignition point for every possible Approximate engine speed to keep constant. For example, 1 ms corresponds to the crankshaft angle of 7.2 ° at 1200 rpm. However, 1 ms corresponds to an engine speed of 3600 Rpm the crankshaft angle of 21.6 °. This causes the fuel injection timing by an angle of about 5 ° an engine speed of 3600 rpm earlier compared to the engine speed of 1200 rpm to have the ignition point at 1200 rpm and 3600 rpm at the same time.

How, however 4 shows, at a low load range in which no smoke is generated, it is not necessary to suppress the generation of smoke, but it is necessary to suppress an abrupt increase in a hydrocarbon HC. For the purpose of eliminating the above problems, the fuel injection timing is shifted to an earlier timing compared to a high load area. This is because if the fuel injection timing is the same as that of high engine load during a low load area in which the wall temperature of the engine combustion chambers is small, an ignition timing is retarded due to an increase in the ignition timing retardation period, so that the combustion chamber temperature would be reduced , which would lead to a high concentration of HC.

Since the fuel injection timing as in 4 is shown, regulates the control unit 31 a time (an amount corresponding to the fuel injection timing) at which the solenoid valve 14 which in 3 is shown is closed.

5 Fig. 14 shows an operation flowchart in which the fuel injection timing and the fuel injection interval (fuel injection amount) are controlled and which is carried out at constant time periods.

According to step 1 5 reads the CPU of the control unit 31 the engine speed Ne, the throttle valve opening angle Acc, the coolant temperature Tw, and the fuel temperature TF. The engine speed Ne is both based on the reference pulse (one pulse per revolution of the fuel pump 20 generated) and a scale impulse ( 36 Pulses per revolution of the fuel pump). Every sensor 34 . 35 serves to detect the coolant temperature Tw and the fuel temperature TF.

According to step 2 5 a fuel base injection timing Itm and fuel base injection interval Avm are determined by the respective ge associated table is read, which contains the engine speed Ne and the throttle valve opening angle Acc. The table (not shown) of the basic fuel injection timing Itm is a table which has the throttle valve opening angle Acc and the engine speed Ne as parameters by which in the diagram after 4 to realize shown fuel injection timing characteristic.

As in 6 is shown, the basic fuel injection interval Avm becomes longer as the throttle valve opening angle Acc increases.

In addition, a fuel injection correction value ΔItm is determined in consideration of the fuel temperature TF and the coolant temperature Tw, so that the fuel injection correction value ΔItm is added to the base fuel injection timing Itm to be according to 5 correct the fuel injection timing in steps 3 and 4.

The fuel injection timing correction value ΔItm is additionally corrected by two correction values ΔItm ₁ and ΔItm ₂ .

7 shows a diagram of the one correction value ΔItm ₁ (fuel temperature correction value), and 8th shows a diagram of the other correction value ΔItm ₂ (coolant temperature correction value).

According to each of the characteristic curves after 7 and 8th a positive angle correction value increases as the corresponding temperature decreases. This is because if the corresponding temperature becomes lower, the rate of combustion will slow down. Temperature compensation is carried out in this way.

In a step 5 according to 5 the derived fuel injection timing IT (= Itm + ΔItm) and the basic fuel injection interval Avm are stored in predetermined addresses.

At the time of injection IT the solenoid valve 14 closed and then the solenoid valve 14 opened at a time when the basic fuel injection interval Avm has passed.

The following is the operation of the preferred embodiment with reference to FIG 9 explained in more detail.

9 Figure 3 shows concentration characteristics of both NO _x and smoke related to the EGR rate in cases where the fuel injection timing is at crankshaft angle before top dead center (BTDC) and where the fuel injection timing is at crankshaft angle behind top dead center (ATDC) lies.

Although the concentration of NO _x decreases with increasing exhaust gas recirculation rate at a fuel injection time before top dead center (IT = -4 ATDC), the concentration of smoke increases abruptly in the form of a progressively increasing curve, as in 9 is shown.

However, if the fuel injection timing is at a crankshaft angle past top dead center (IT = + 4 ATDC), there is a tendency for the smoke to decrease as well as NO _x with increasing exhaust gas recirculation rate.

The reasons for a decrease in smoke concentration are as shown in the heat outlet diagram 9 It is shown that due to the extremely large delay in the injection timing and the higher exhaust gas recirculation rate, the ignition timing delay period is significantly extended, so that a high percentage of the combustion takes place in the premix-air combustion. In other words, if the injection timing is conventionally retarded to a crank angle behind top dead center when the exhaust gas recirculation rate is not as high, the tendency to smoke concentration higher than in FIG 10 shown are not suppressed.

Because according to the preferred embodiment a large Percentage of combustion taking place in the premix air combustion the smoke concentration can be significantly reduced, even if the operation takes place at a high exhaust gas recirculation rate.

According to 9 For example, the crankshaft angle behind top dead center is 4 °. Because the critical point for both premix air combustion and diffusion combustion differs depending on the engine used, adjustment to the particular engine is required, with the specific crankshaft angle behind top dead center for each of these engines must be determined.

It also shows 11 a characteristic of the fuel consumption rate for this first embodiment.

Because of the delay in fuel injection timing according to the first embodiment a degree of constant displacement of the diesel engine is deteriorated, but a cooling loss significantly reduced due to the reduction in the combustion temperature.

Because of this, deteriorates the fuel consumption rate, although a delay in the Fuel injection timing takes place. The degree means the constant volume a work efficiency and this work efficiency is defined as the following value: work efficiency = indexed Work / heat generated = indicated power / (1 - cooling loss).

In the first embodiment, the temperature compensation is compensated for by a more advanced angle of the delayed injection timing when both the fuel temperature and the cooling temperature become lower, so that the same ignition timing at both the lower temperature time and the higher temperature temperature can be reached.

12 shows a second embodiment.

According to the second embodiment is a pre-compression device 41 provided a mechanical pre-compressor 42 has and is provided with an adjustable belt drive and a device for controlling a swirling device speed in addition to the exhaust gas recirculation system, as is already provided in the first embodiment.

The pre-compressor 42 is in the inlet air duct 23 arranged behind the mixing point of the exhaust gas recirculation gas with the intake air and with the engine 21 by means of a pulley 45 , which in turn is coupled to the crankshaft of the engine, and by means of an adjustable pulley 43 and a belt wrapped around the two pulleys 44 connected. With the adjustable pulley 43 the pulley profile by means of an actuating device (not shown) as a function of a signal from the control unit 51 adjusted so that a speed ratio between the engine 21 and the pre-compressor 42 gets bigger or smaller.

The control unit 51 regulates the speed ratio described above so that an approximately constant pre-compression pressure of 400-500 mmHg is maintained over the entire engine speed range.

13 shows a course of the speed ratio.

If the engine speed is relatively low, the speed ratio is 3: 1, so that the speed of the pre-compressor 42 is increased so that the precompression pressure increases. With the high speed range of the engine speed, the exhaust gas recirculation rate becomes lower, the concentration of HC is reduced, and the maximum internal cylinder pressure (Pmax) is increased accordingly. At this time, the speed ratio is small and is 1: 1, so that the pre-compression pressure is not increased.

In addition, in the second embodiment, when the intake air with the exhaust gas mixture by means of the exhaust gas recirculation channel 26 in the pre-compressor 42 is generated, a deposit due to the carbon component or soot component of the exhaust gas is generated.

In order to prevent such deposits, a screw-type pre-compressor is used 42 used, in which the rigidity of the blades is high and line-to-line contact with the housing is realized.

In addition, a rotor blade vortex control device is in the 14 and 15 shown.

• a) einem Rotorblatt 47, welches drehbar an einer Stelle in der Nähe eines Wirbelkanals 46b eines sogenannten Spiral-Einlaßtors 46 installiert ist (welches mit einem im wesentlichen geradlinigen Einlaßkanal 46a und einem Wirbelkanal 46b versehen ist, welcher um eine Achse des Einlaßventils des Motors gewunden ist);
• b) einem Verbindungsmechanismus 49, welcher mit dem Blatt 47 verbunden ist; und
• c) einer Betätigungseinrichtung 48 zum Betätigen der Verbindungseinrichtung 49.

The rotor blade vortex control device is provided with:

• a) a rotor blade 47 which is rotatable at a location near a spinal canal 46b a so-called spiral inlet gate 46 installed (which has an essentially straight inlet duct 46a and a spinal canal 46b is provided, which is wound around an axis of the intake valve of the engine);
• b) a link mechanism 49 which with the sheet 47 connected is; and
• c) an actuator 48 to operate the connection device 49 ,

The setting of the swirl ratio is in a rotational position of the rotor blade 47 possible. For example, the sheet position after 14 achieved a high swirl ratio. However, if the sheet 47 in the 15 shown position reached, a low swirl ratio is achieved. This rotor blade vortex control device responds very quickly and a large (continuous or gradual) area of the vortex control can be reached. A suitable HC control is therefore possible, which reacts strongly to the swirl ratio.

16 shows a diagram of the swirl ratio related to the engine operating conditions.

How 16 shows, the swirl ratio increases with decreasing engine speed. In the high-speed range, a reduction in volume efficiency goes hand in hand with a high swirl ratio. The improvement of the combustion due to a high pressure injection requires the need to weaken the swirl so that the swirl ratio is gradually reduced as the engine speed increases.

The adjustable swirl ratio actuator 48 has a diaphragm actuator with a two-stage spring (not shown) and a vacuum control valve, which in three stages generates a controlled vacuum by diluting the atmosphere to a constant vacuum, which comes from a vacuum source.

The control unit 51 generates a base swirl ratio by reading a table (not shown) of the swirl ratio (base swirl ratio), the table containing the engine speed Ne and the throttle valve opening angle Acc as parameters. The opening angle Vb of the vacuum control valve according to the read base swirl ratio is read and stored in a predetermined address (steps 11-14 in 17 ).

18 shows respective concentration characteristics of the concentration of HC, smoke, and NO _x based on the exhaust gas recirculation rate.

The symbols R, R + C and R + C + S show in 18 following:
R: the case where the fuel injection timing is at a crank angle behind the top dead point is delayed (corresponds to the first embodiment);
R + C: the case in which precompaction takes place in addition to R;
R + C + S: the case in which precompression and vortex control take place in addition to R.

This is enough for the aforementioned embodiment the pre-compression pressure was 400-500 mmhg and the swirl ratio 5th

In the first embodiment, the concentration of HC was not described. The combination of a large exhaust gas recirculation rate and an extreme delay in the injection timing causes the fuel temperature to decrease, so that although the NO _x and smoke concentrations can be reduced significantly, in absolute terms there is a lack of oxygen and, as a result, the concentration of HC tends to increase, as shown by the Curve "R" in 18 is shown. To suppress an increase in the concentration of HC, an oxidizing agent is required 40 be provided, which is usually in the exhaust duct 25 according to the first, in 1 shown embodiment is arranged.

According to the second embodiment, as in 18 It is shown that the concentration of HC as well as the concentration of NO _x and smoke are significantly reduced in an engine operating range in which the high exhaust gas recirculation rate is used. In other words, according to the second embodiment, it is not necessary to use an oxidizing agent 40 to be provided in the exhaust gas duct as required in the first exemplary embodiment.

In addition to the pre-compression in the low speed range of the engine, the absolute amount of oxygen required can be ensured, and an increase in the vortex produces an improvement in fuel combustion, so that the concentration of HC can be significantly reduced to a level at which a unit degree of exhaust gas is reduced to one predetermined limited value can be set without an oxidizing medium 40 to use.

In 19 is a third preferred embodiment.

According to the third embodiment, an oxygen removal filter (a so-called oxygen deficiency film) is provided, as will be described later, in order to reduce the oxygen concentration of the intake air, instead of inert exhaust gas recirculation gases into the intake duct for this purpose 23 by means of the exhaust gas recirculation system so as to reduce the oxygen concentration of the intake air in another way.

How 19 shows, the inlet air duct branches 23 in two channels on one downstream of the mechanical pre-compressor 42 located, and an oxygen removal filter 63 is in a branch duct 61 arranged. The oxygen removal filter 63 is known and although oxygen can pass through it, nitrogen cannot pass through it, using hollow channels which are adapted to the size of the oxygen molecules or nitrogen molecules.

The one with the filter 63 removed oxygen is through the exhaust duct 25 by means of a connecting channel 64 led out.

A flow control valve 65 is arranged in the branched part to the intake air flow amount in the branch duct 61 adjust in which of the filters 63 is arranged. The flow control valve 65 is formed by a throttle valve (not shown) and an electromagnetic valve, which the opening position of the throttle valve in three stages corresponding to one of the control unit 71 output signal can adjust.

When the throttle valve is in its fully closed position, the total amount of intake air is in the branch passage 61 passed and with increasing opening angle of the throttle valve is in the branch duct 61 headed intake air volume reduced. In other words, when an intake air amount of 10% in the branch duct 61 flows, the oxygen concentration in the intake air is 15%, or is further increased when the intake air flow through the branch duct 62 is directed, is increased. In this way, the concentration of oxygen in the intake air is gradually increased to 17% and 19% (note that the air-oxygen concentration is 21%).

The oxygen concentration set in the three stages is based on the operating conditions of the engine in 20 shown.

The stepwise characteristic of the oxygen concentrations corresponds to the stepwise characteristic of the in 2 Exhaust gas recirculation rate shown. The oxygen concentration levels 15%, 17% and 19% correspond to the associated exhaust gas recirculation rates of 100%, 60% and 30%, respectively.

For this reason, the oxygen removal filter film is provided in such an amount that a reduction in the oxygen concentration up to 15% can be set at a pre-compression pressure of 400 mmhg. The constant control of the pre-compression pressure is exerted such that an inlet pressure of the oxygen removal filter 63 when using the pre-compressor 42 has an amount of 400 mmhg.

In the third embodiment, the oxygen removal filter works 63 and the flow control valve 65 in the same way as exhaust gas recirculation system according to the second embodiment described above.

Due to the functional differences of the filter 62 and the control valve 65 related to the exhaust gas recirculation system eliminates the need to return the exhaust gas to the intake duct. As a result, contamination due to soot separation from the exhaust gas cannot occur in the exhaust valve and the pre-compressor, and moreover, the increase in the NO _x concentration which might otherwise occur according to an increase in the intake air temperature due to the high exhaust gas temperature can be suppressed.

To lower the combustion temperature, reducing the engine compression ratio and / or pre-cooling the intake air as well as reducing the oxygen concentration of the intake air using the exhaust gas recirculation system according to the first and second embodiments and the oxygen removal filter and the control valves 63 and 65 according to the third embodiment.

Concentration without an increase in - the method as described above for controlling the combustion of the diesel engine, the combustion temperature of the diesel engine according to the engine operating conditions is reduced and the ignition delay period greatly extended in the engine operating region in which the combustion temperature is low, the NO _x can the smoke concentration can be reduced.

Different effects can be caused by the described embodiments so alike be achieved.

Claims (6)

Method for controlling a diesel engine, comprising the steps: detecting an engine operating state, controlling a fuel injection in accordance with the detected operating state of the engine, returning an exhaust gas via an exhaust gas recirculation valve between an inlet duct and an exhaust gas duct, detecting an exhaust gas recirculation ratio (EGR ratio), characterized in that the exhaust gas recirculation ratio (EGR ratio) is controlled as a function of the detected engine operating state and at the same time a fuel injection time (IT) of a fuel injection valve ( 20 ) depending on the detected exhaust gas recirculation ratio (EGR ratio). Method for controlling a diesel engine according to claim 1, characterized in that with a comparatively high exhaust gas recirculation ratio (EGR ratio) the fuel injection time (IT) of the fuel injection valve ( 20 ) to a late crank angle that includes a crank angle (BTDC) approximately before an upper adjacent top dead center position (TDC), top dead center (TDC) or crank angle (ATDC) after top dead center (TDC) of a piston stroke in each cylinder is set by a fuel injection control device to extend an ignition delay of the fuel in each combustion chamber of the engine ( 21 ). Method for controlling a diesel engine according to claim 1 or 2, characterized in that an oxygen concentration in a in the engine ( 21 ) inlet air supplied is reduced. Method for controlling a diesel engine according to claims 1 to 3, characterized in that an intake air pre-compression by means of a pre-compressor ( 41 . 42 ) occurs when the fuel injection timing (IT) is decelerated to a crank angle behind top dead center (TDC). Method for controlling a diesel engine according to claims 1 to 4, characterized in that a swirl of the intake air is by means of a swirl control device ( 46 - 49 . 51 ) is amplified when the deceleration of the fuel injection timing (IT) is carried out to an injection timing after the top dead center (TDC) of the piston stroke. Method for controlling a diesel engine according to claim 3, characterized in that the inlet air through a branch pipe ( 61 ) flows from at least two branch ducts and the oxygen concentration of the inlet air by means of an oxygen removal filter ( 63 ) is reduced, as well as the inlet air in the direction of a branch duct by means of a flow quantity control valve ( 65 ) which is regulated according to the engine operating state.