JPH11315739A - Diesel engine combustion control system - Google Patents

Abstract

PROBLEM TO BE SOLVED: To permit low temperature pre-mixed combustion in the low load region with a low compression ratio setting, which is difficult to realize low temperature pre-mixed combustion. SOLUTION: In this diesel engine, exhaust is circulated, engine compression ratio is set low, and pre-mixed combustion is performed at low temperatures. When the atmospheric temperature at combustion start, which is determined according to the compression ratio and intake gas temperature, is judged in a region lower than a first target temperature for maintaining pre-mixed combustion, a control means 85 activates a temperature rise control unit 82 for increasing the operation gas temperature in the cylinder such that the temperature at combustion start exceeds the first target temperature. Furthermore, fuel injection timing of a fuel injection valve 81 is controlled to be within the range between 5 deg. and 20 deg. after the compression top dead center. Accordingly, the low temperature pre-mixed combustion is stabilized in the low load region where combustion tends to be unstable with a low compression ratio.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

The present invention relates to a combustion control device for a diesel engine.

[0002]

2. Description of the Related Art Diesel engine combustion methods include:
There is low-temperature premixed combustion that reduces NOx and particulates emitted from the engine. This delays the fuel injection timing until after the compression top dead center, and also causes the exhaust gas recirculation (E
By reducing the oxygen concentration by GR), the ignition delay period of the fuel is prolonged, and during this ignition delay period, a premixed gas in which the fuel is sufficiently vaporized is formed, and low-temperature premixed combustion is performed with low-concentration oxygen.

[0003] In this case, there is a proposal that the low-temperature premixed combustion can be performed even on the high-load side of the engine by setting the engine compression ratio to be lower than the normal compression ratio (Japanese Patent Laid-Open No. 8-254134). reference).

[0004] When the compression ratio of the engine is high, the compression temperature relatively rises, so that the temperature at the start of combustion (the ambient temperature in the cylinder at the start of combustion) increases. Since the combustion temperature further rises after the start of combustion, if it exceeds the range of low-temperature premix combustion, low-temperature premix combustion cannot be performed.

[0005] In particular, when the engine load increases, the temperature of the intake gas rises because the temperature of the exhaust gas recirculated in the intake air rises, and the actual compression ratio increases due to the supercharging of the intake air by the supercharger. The temperature rises beyond the temperature range of the low-temperature premixed combustion, so that the low-temperature premixed combustion can be performed only on the low load side.

[0006] On the other hand, by lowering the compression ratio by the proposed device, the combustion start temperature is relatively lowered, and low-temperature premixed combustion can be performed on a higher load side.

[0007]

However, if the compression ratio is reduced, the combustion start temperature in the low load region will be too low, and the combustion will be unstable on the low load side.

Therefore, in the above-described conventional apparatus, on the low load side where low-temperature premixed combustion cannot be performed, the oxygen concentration is increased by decreasing the EGR amount, and the fuel injection timing is advanced, so that the combustion start is advanced. To stabilize combustion.

However, in this case, the amount of NOx emission increases due to the decrease in the EGR amount and the advance of the injection timing, and the low-temperature premixed combustion is not performed, so that the combustion noise increases and the fuel efficiency deteriorates. .

Accordingly, an object of the present invention is to enable low-temperature premix combustion in a wide operating range of an engine even when the compression ratio is set lower than a normal compression ratio.

[0011]

According to a first aspect of the present invention, there is provided a diesel engine in which low-temperature premix combustion is performed at a low compression ratio, as shown in FIG. A temperature raising control device 82 for raising the temperature of the working gas in the combustion chamber, a means 83 for predicting the in-cylinder ambient temperature at the start of the combustion, and Means 84 for determining whether or not the temperature is lower than the first target temperature T1
When the temperature in the cylinder at the start of combustion is determined to be lower than the first target temperature T1, the temperature rise control device 82 is controlled so that the temperature in the cylinder at the start of combustion exceeds the first target temperature. Means 85 for controlling the temperature rise by operating the fuel injection valve 81 and adjusting the fuel injection timing of the fuel injection valve 81 so that the rate of increase in the combustion temperature becomes a predetermined value or more.

In a second aspect, in the first aspect, the cylinder ambient temperature t2 at the time of the start of combustion is predicted based on the compression ratio ε and the engine load.

According to a third aspect, in the first aspect, the cylinder ambient temperature t2 at the start of the combustion is predicted based on the compression ratio ε and the intake gas temperature.

According to a fourth aspect of the present invention, in any one of the first to third aspects of the present invention, the combustion is started 5 ° after the compression top dead center so that the temperature rise rate of the main combustion is equal to or higher than the predetermined value. The fuel injection timing is controlled so as to fall within the range of ゜ 20 °.

According to a fifth aspect of the present invention, in any one of the first to fourth aspects of the present invention, an intake valve closing timing control mechanism is provided as the temperature raising control device, and the cylinder ambient temperature t2 at the start of combustion is reduced by the second temperature. When the temperature is equal to or lower than the first target temperature T1 and higher than the second target temperature T2 lower than the first target temperature T1, the closing timing of the intake valve is advanced,
The temperature raising control is performed.

According to a sixth aspect of the present invention, in the fifth aspect, a swirl control mechanism is further provided as the temperature raising control device, and an in-cylinder ambient temperature t2 at the start of the combustion is equal to or lower than the second target temperature T2, and When the temperature is higher than the third target temperature T3 lower than the second target temperature T2, the temperature increase control is performed by advancing the intake valve closing timing and lowering the swirl ratio.

According to a seventh aspect of the present invention, in the fifth aspect, a supercharging pressure control mechanism is further provided as the temperature raising control device, wherein the in-cylinder ambient temperature t2 at the start of the combustion is equal to or lower than the second target temperature T2, and When the temperature is higher than the third target temperature T3, which is lower than the second target temperature T2, the intake valve closing timing is advanced, the swirl ratio is reduced, and the supercharging pressure is further increased. Do.

According to an eighth aspect of the present invention, in the sixth or seventh aspect, a fuel injection control mechanism is further provided as the temperature raising control device, and the cylinder internal temperature t2 at the start of the combustion is lower than the third target temperature T3. And the third target temperature T
When the temperature is higher than the fourth target temperature T4, which is lower than 3, the intake valve closing timing is advanced, the swirl ratio is lowered, the pilot injection is performed prior to the main fuel injection, and the combustion by the pilot injection is performed mainly. The temperature increase control is performed by setting the pilot injection timing so as to end before the injection.

According to a ninth aspect, in the sixth or seventh aspect, a fuel injection control mechanism is further provided as the temperature increase control device, and the in-cylinder ambient temperature t2 at the start of the combustion is lower than the third target temperature T3. And the third target temperature T
When the temperature is higher than the fourth target temperature T4, which is lower than 3, the intake valve closing timing is advanced, the swirl ratio is reduced, the supercharging pressure is increased, and the pilot injection is performed prior to the main fuel injection. And by setting the pilot injection timing so that the combustion by the pilot injection ends before the main injection,
The temperature raising control is performed.

According to a tenth aspect, in the eighth or ninth aspect, the main fuel injection timing is controlled so that the main injected fuel is ignited after the end of the main fuel injection.

According to an eleventh aspect of the present invention, in any one of the eighth to tenth aspects of the present invention, the opening / closing timing of one of the plurality of exhaust valves provided for one cylinder as the temperature rise control device is arbitrarily determined. And an EGR valve for controlling the amount of exhaust gas recirculated during intake. When the in-cylinder ambient temperature t2 at the start of combustion is equal to or lower than the fourth target temperature T4, the intake valve is closed. Advance the timing, lower the swirl ratio, increase the supercharging pressure, perform pilot injection so that main injection is performed after combustion of pilot injection is completed, and open the one exhaust valve in the intake stroke At the same time, the temperature increase control is performed by closing the EGR valve.

According to a twelfth aspect of the present invention, in any one of the eighth to tenth aspects of the present invention, the opening / closing timing of one of the plurality of exhaust valves provided for one cylinder as the temperature rise control device is arbitrarily determined. Exhaust valve opening and closing mechanism that can be controlled
An intake valve control mechanism that can arbitrarily adjust the opening / closing timing of the intake valve; and an EGR valve that controls the amount of exhaust gas recirculated during intake, wherein the in-cylinder ambient temperature t2 at the start of combustion is the fourth target temperature. When T4 or less, the intake valve closing timing is advanced, the swirl ratio is reduced, the supercharging pressure is increased, pilot injection is performed so that main injection is performed after combustion of pilot injection is completed, and the EGR is performed. The respective amounts of the EGR gas and the intake air flowing into the combustion chamber via the valve are reduced as the cooling water temperature is lowered, and the amount of the EGR gas flowing back into the combustion chamber by opening one exhaust valve in the intake stroke is reduced by the cooling water temperature. The temperature rise control is performed by increasing as the value becomes lower.

According to a thirteenth aspect, in the eleventh or twelfth aspect, the exhaust port on the one exhaust valve side which is opened in the intake stroke is formed as a helical port.

According to a fourteenth aspect, in any one of the first to thirteenth aspects, the target temperatures T1, T2,
T3 and t4 are set according to the load such that the lower the engine load, the lower the temperature.

A fifteenth aspect of the present invention relates to a fuel injection valve 81 in which a fuel injection timing is variable as shown in FIG.
And a temperature raising control device 82 for increasing the temperature of the working gas in the combustion chamber, a means 91 for detecting an engine load that affects the ambient temperature in the cylinder at the start of combustion, and the engine load maintains low-temperature premixed combustion. Means 92 for determining whether or not the vehicle is in a load region (region C) which is lower than or equal to the first target load and higher than the second target load which is lower than the first target load.
When it is determined that the engine load is equal to or less than the first target load and is higher than the second target load, the temperature raising control device 82 is operated to perform the temperature raising control, and the combustion temperature is reduced. Means 93 is provided for adjusting the fuel injection timing of the fuel injection valve 81 so that the rate of increase is equal to or higher than a predetermined value.

According to a sixteenth aspect, in the fifteenth aspect, a swirl control mechanism is further provided as the temperature raising control device, and a swirl control mechanism is provided which is lower than the second target load and lower than the third target load. In a high load region (D region), the intake valve closing timing is advanced, and the swirl ratio is lowered to perform the temperature increase control.

In a seventeenth aspect, in the fifteenth aspect, a supercharging pressure control mechanism is further provided as the temperature raising control device,
In a load region (D region) that is equal to or less than the second target load and higher than the third target load that is lower than the second target load,
The temperature increase control is performed by advancing the intake valve closing timing, lowering the swirl ratio, and further increasing the supercharging pressure.

According to an eighteenth aspect, in the fifteenth or sixteenth aspect, a fuel injection control mechanism is further provided as the temperature increase control device, and the fuel injection control mechanism is provided at the third target load or less and the third target load or less.
A load range lower than the target load and higher than the fourth target load (E
Region), the intake valve closing timing is advanced, the swirl ratio is lowered, pilot injection is performed prior to main injection of fuel, and combustion by pilot injection is completed so as to end before main injection. The temperature increase control is performed by setting the injection timing.

According to a nineteenth aspect, in the fifteenth or sixteenth aspect, the temperature increase control device further includes a fuel injection control mechanism, and the fuel injection control mechanism is provided at a pressure equal to or less than the third target load.
A load range lower than the target load and higher than the fourth target load (E
Region), the intake valve closing timing is advanced, the swirl ratio is reduced, the supercharging pressure is increased, and the pilot injection is performed prior to the main fuel injection. The temperature increase control is performed by setting the pilot injection timing so as to end before.

In a twentieth aspect, in the eighteenth or nineteenth aspect, the main fuel injection timing is controlled so that the main injected fuel is ignited after the end of the main fuel injection.

[0031]

According to the first, second, third, fourth, and fourteenth aspects of the present invention, the atmospheric temperature in the cylinder at the start of combustion becomes lower than the first target temperature for maintaining low-temperature premixed combustion, so that it remains unchanged. In the temperature range where low-temperature premixed combustion cannot be performed, the temperature control is performed by operating the temperature raising control device so that the ambient temperature in the cylinder at the start of combustion exceeds the first target temperature. The fuel injection timing of the fuel injection valve is adjusted so as to be equal to or more than a predetermined value, so that the low-temperature premix combustion is performed at a low temperature or a low load region where the low-temperature premix combustion cannot be performed according to the conventional device. (Expansion of the low-temperature premixed combustion zone), which results in combustion noise,
All of the fuel efficiency, HC, PM, and NOx could be improved compared to the conventional device.

Further, since the low-temperature premixed combustion can be realized at low temperatures and on the low-load side where it was difficult to perform the low-temperature premixed combustion according to the conventional apparatus, the compression can be further reduced as compared with the conventional apparatus. Thus, the low-temperature premixed combustion region can be shifted to a higher load side.

In the fifth and fifteenth aspects of the invention, the actual compression ratio increases with the advance of the intake valve closing timing. According to the fifth aspect of the present invention, the cylinder ambient temperature at the start of combustion is lower than the first target temperature. Also, when the temperature is higher than the second target temperature, and according to the fifteenth aspect, also in a load region (C region) that is equal to or less than the first target load and higher than the second target load, Drop can be prevented.

In each of the sixth, seventh, sixteenth, and seventeenth inventions, the swirl is weakened, so that the cooling loss during the compression stroke is reduced. When the cylinder ambient temperature is equal to or lower than the second target temperature and higher than the third target temperature,
According to the seventh aspect of the invention, it is possible to prevent a decrease in the combustion start temperature even in a load region (D region) that is equal to or lower than the second target load and higher than the third target load.

However, in this state, combustion deterioration occurs due to a decrease in the swirl ratio. However, according to the seventh and seventeenth aspects, the supercharging pressure is increased to increase the amount of intake air, thereby decreasing the swirl ratio. By increasing the working gas to maintain the angular kinetic energy at the compression top dead center, combustion deterioration can be prevented, and this high supercharging raises the intake air temperature, contributing to the rise in combustion start temperature. I do.

In the eighth, ninth, tenth, eighteenth and nineteenth aspects of the present invention, the pilot injection is performed.
According to the invention, when the in-cylinder ambient temperature at the start of main combustion combustion is equal to or lower than the third target temperature and higher than the fourth target temperature, in the eighteenth and nineteenth inventions, it is equal to or lower than the third target load. In addition, even in a load region (E region) higher than the fourth target load, it is possible to increase the in-cylinder ambient temperature at the start of main combustion combustion.

According to the eleventh and twelfth aspects, low-temperature premixed combustion can be realized even when the in-cylinder ambient temperature at the start of combustion is equal to or lower than the fourth target temperature. Is further improved.

According to the thirteenth aspect, it is possible to prevent a reduction in the swirl ratio in the combustion chamber due to the opening of one of the exhaust valves in the intake stroke.

When the region determination is performed by the engine load, the accuracy of the region determination is reduced when the intake air temperature is different from the matching temperature. However, in the fifth to fourteenth aspects, the combustion start temperature Because the area is determined based on the difference between the summer and winter, regardless of the intake air temperature,
Region determination can be performed with high accuracy.

[0040]

FIG. 1 is a schematic diagram of a diesel engine.

In the combustion of a diesel engine, NO
The generation amount of x greatly depends on the combustion temperature, and it is effective to lower the combustion temperature relatively to reduce it. In the low temperature premix combustion system, the oxygen concentration is reduced by an exhaust gas recirculation system (EGR), thereby realizing low temperature combustion. For this reason, the exhaust passage 2 and the intake passage 3 are connected by an EGR passage 4, and a diaphragm type EGR valve 6 that operates according to the control negative pressure from the negative pressure control valve 5 is provided in the EGR passage 4. Part of the exhaust gas is recirculated into the intake air.

The negative pressure control valve 5 includes a control unit 4
Driven by a duty control signal from 1
An appropriate EGR rate is obtained according to the operating conditions of the engine. For example, the EGR rate is set to the maximum 100% (the intake air flow rate and the EGR gas flow rate are equal to each other) in the low rotation and low load range, and the EGR rate increases as the rotation speed and the load increase.
Decrease rate. Since the exhaust gas temperature rises on the high load side, if a large amount of EGR gas is recirculated, the intake air temperature rises,
As a result, the combustion temperature also rises relatively, the effect of reducing NOx decreases, and the ignition delay period of the injected fuel becomes short, so that premixed combustion cannot be realized. For this reason EG
The R ratio is reduced as the load becomes higher.

An EGR gas cooling device 7 is provided in the EGR passage 4. It has a water jacket 8 formed around the EGR passage 4, in which a part of the engine cooling water is circulated, and the amount of the cooling water circulated depends on the flow rate provided in the cooling water inlet 7 a. It can be adjusted by the control valve 9. As the degree of opening of the control valve 9 increases according to a command from the control unit 41, the degree of cooling of the EGR gas increases.

A swirl control valve (not shown) is provided in the intake passage near the intake port of the engine. The opening of the swirl control valve is controlled by the control unit 41, and when the valve is closed in a low engine speed and low load range (the opening is reduced), the flow velocity of the intake air sucked into the combustion chamber is increased and the swirl is strong in the combustion chamber. Is generated. However, when the swirl increases, the heat exchange rate of the working gas in the cylinder increases, and the working gas temperature relatively decreases.

A hollow combustion chamber (not shown) formed in the piston is a large-diameter toroidal combustion chamber. In this, the piston cavity is formed in a cylindrical shape from the crown to the bottom of the piston without narrowing the inlet.A cone is formed in the center of the bottom, and this cone allows the piston to enter the piston cavity at the latter stage of the compression stroke. Further, the mixing of air and fuel is further improved so as not to give resistance to swirl flowing while turning.

As described above, the swirl generated by the above-described swirl control valve due to the cylindrical piston cavity which does not restrict the inlet moves from inside the piston cavity to outside the cavity as the piston descends in the combustion process. It is diffused and the swirl is maintained outside the cavity.

The exhaust passage 2 is provided with a turbocharger downstream of the branch point of the EGR passage 4. This turbocharger is provided at the scroll inlet of the exhaust turbine 52,
A variable vane 53 driven by a step motor 54 is provided. The variable vane 53 is controlled by the control unit 41, and the exhaust turbine 52 is provided on the low rotation speed side so that a predetermined supercharging pressure is obtained from the low engine speed range.
Is controlled to a vane angle that increases the flow velocity of the exhaust gas introduced into the exhaust turbine 52, and is controlled to a vane angle (fully open state) where the exhaust gas is introduced into the exhaust turbine 52 without resistance on the high rotation side. Also,
Depending on the operating conditions, the variable vane 53 is controlled to a vane angle at which a desired boost pressure is obtained.

The engine is provided with a common rail type fuel injection device 10.

This mainly includes a fuel tank (not shown), a supply pump 14, a common rail (accumulator) 1
6. A fuel injection nozzle 17 provided for each cylinder. The high-pressure fuel generated by the high-pressure supply pump 14 is stored in the common rail 16, and the three-way valve 25 in the fuel injection nozzle 17 opens and closes the nozzle needle to inject the fuel. Start and end can be controlled freely. The fuel pressure in the common rail 16 is always controlled to an optimum value required by the engine by a pressure sensor (not shown) and a discharge amount control mechanism (not shown) of the supply pump 14.

The control of the fuel injection amount, injection timing, fuel pressure, etc. is performed by a control unit 41 composed of a microprocessor. Therefore, the control unit 41 receives signals from an accelerator opening sensor 33, a sensor 34 for detecting an engine speed and a crank angle, a sensor 35 for determining a cylinder, and a water temperature sensor 36. Control unit 41
Calculates the target fuel injection amount and the fuel injection timing according to the engine speed and the accelerator opening, controls the on-time of the three-way valve 25 in the nozzle corresponding to the target fuel injection amount, The ON timing of the three-way valve 25 is controlled according to the injection timing. Further, the fuel pressure of the common rail 16 is feedback-controlled via the discharge amount control mechanism of the supply pump 14 so that the common rail pressure detected by a pressure sensor (not shown) matches the target pressure.

The fuel injection timing is retarded from the normal injection timing in order to realize low-temperature premix combustion. As described later, the fuel injection is set to start within a predetermined range after the compression top dead center at the crank angle. As a result, the ignition delay period of the injected fuel is prolonged, and during this time, the vaporization of the fuel is promoted, and the ignition can be performed in a state where the fuel is sufficiently mixed with the air. Thereby, low-temperature premix combustion is performed under a low oxygen concentration due to exhaust gas recirculation, and NOx can be reduced without increasing particulates.

Here, the combustion region of the low-temperature premixed combustion will be described. In FIG. 2, the shaded region is a region where low-temperature premix combustion is possible (the lower limit is, for example, about 600 ° C., and the upper limit is, for example, about 700 ° C.). Even if the temperature is higher or lower than this range, low-temperature premix combustion cannot be performed.

If the compression ratio of the engine is high, the combustion start temperature (atmospheric temperature at the start of combustion) becomes relatively high.
At high loads, the combustion temperature becomes too high, and low-temperature premixed combustion can be performed only at low loads.

At a high load, the temperature of the exhaust gas recirculated during the intake increases, the intake gas temperature increases, and the actual compression ratio increases due to the supercharging of the intake. As a result, even at the same compression ratio, the gas temperature in the cylinder at the time of the maximum compression increases, and the combustion start temperature increases.

By the way, the above-mentioned Japanese Patent Application Laid-Open No. 8-25413 is disclosed.
In the conventional device disclosed in Japanese Patent Publication No. 4 (2004), it is possible to shift the low-temperature premixed combustion region to a higher load side by lowering the compression ratio and thereby lowering the combustion start temperature.

On the other hand, on the other hand, the combustion start temperature in the low load region becomes too low, and the low temperature premixed combustion becomes very unstable on the low load side.

In order to enable low-temperature premixed combustion on the low load side, in the present invention, the compression ratio of the engine is set to a compression ratio of 16 or less, which is lower than that of a normal engine. By operating a temperature raising control device to be described later so that the combustion temperature becomes relatively high, the temperature raising control is performed, and the fuel injection timing is set to, for example, 5 ° to 20 ° after the top dead center.
Is set to increase the temperature rise rate after the start of combustion.

This will be further described with reference to FIG. 3. The combustion start temperature t2 is in the illustrated premixed combustible zone (the temperature range from the target temperature T0 to the target temperature T1 (where T1 <T0)), and When the start of combustion is a crank angle and 5 ° to 20 ° after compression top dead center (experimental value), the target temperature T ′ is always reached after the start of combustion, and in this case, low-temperature premixed combustion may be performed. Has been confirmed by experiments.

The target temperatures T0 and T1, which determine the upper and lower limits of the premixable combustion zone, and the target temperature T '.
Is a value uniquely determined from the fuel amount of the main injection and the main injection timing.

On the other hand, when the combustion start temperature t2 is equal to or lower than the target temperature T1 required for premix combustion, the target temperature T
'Cannot be reached, and no low-temperature premixed combustion is performed.

Therefore, in the present invention, when the combustion start temperature t2 is equal to or lower than the target temperature T1, as described later with reference to FIG. 10, the target temperatures T2, T3, T4 (where T1> T2)
>T3> T4), and the temperature rise control is performed in accordance with these temperature ranges, so that the combustion start temperature is raised to the premixed combustion possible zone.

Next, the main combustion starts at 5 ° after compression top dead center.
It must be 20 ° for the following reasons.

When the start of the main combustion is immediately after the compression top dead center, the temperature in the combustion chamber due to the combustion tends to rise rapidly as shown by the solid line in FIG. However, in practice, the compression temperature decreases as it advances from the compression top dead center, so that the temperature rise when the start of main combustion is after the top dead center, as a result, as indicated by the dashed line, It becomes a gentle rising curve. Therefore, if the target temperature T 'is reached by the temperature rise including the decrease in the compression temperature, the premixed combustion can be performed. However, when the main combustion is started after 20 ° after the compression top dead center, the compression temperature is reduced. At this time, it is not possible to increase to the target temperature T 'as shown by the dotted line. Therefore, the limit on the late side of the start of the main combustion is set to 20 ° after the compression top dead center.

FIG. 5 shows the difference in the heat release rate depending on the combustion start timing. From this figure, when the start of combustion is 20 ° or more after the compression top dead center, the heat release rate pattern becomes as shown by the broken line, and the specific heat release rate pattern at the time of the low-temperature premixed combustion shown by the dashed line. It turns out that it cannot be obtained.

As described above, the fact that the rate of temperature rise from the start of combustion is equal to or higher than the predetermined value and the fact that the combustion start temperature t2 exceeds the target temperature T1 as described above causes low-temperature premixed combustion to be performed. These conditions have been confirmed by experiments.

Next, the combustion control region will be described with reference to FIG.

This figure shows a control region for combustion with respect to the engine load (engine torque) and the number of revolutions. The combustion start temperature t2 is related to the load and the number of revolutions of the engine.

In the present invention, the compression ratio is set as described above.
By making the compression ratio lower than that of the conventional device (the compression ratio of the conventional device is set to 18 or less in the present invention, the compression ratio is set to 16 or less), so that the low-temperature premixed combustion region is further improved. It is expanded to the high load side.

For this reason, on the low load side, the combustion start temperature becomes too low, and the temperature falls outside the low temperature premixed combustion region.
Therefore, in order to raise the combustion temperature, first, the intake valve closing timing is advanced to near the intake bottom dead center to increase the actual compression ratio. If it is not enough to just advance the intake valve closing timing, in order to further raise the temperature, the actual compression ratio is further increased by lowering the swirl ratio and increasing the supercharging pressure. At the load, pilot injection is further performed to perform preliminary combustion, thereby increasing the combustion start temperature to enable low-temperature premixed combustion.

The control in each area shown in FIG.
It will be described in detail with reference to FIG.

<1> Region A Since EGR is not performed at the maximum load point, the oxygen concentration is high, and the supercharging pressure is high, so that the combustion start temperature is high, and so-called ordinary diesel combustion is performed.

<2> Region B This region is a low-temperature premixed combustion region in which the load is smaller than the maximum load region, in which the oxygen concentration is reduced by a large amount of EGR, and the temperature of the EGR gas is reduced by the cooling device 7. To control the intake air temperature, and furthermore, the combustion start temperature is lowered by delaying the main injection timing and reducing the supercharging pressure. As a result, the ignition delay period is prolonged, the fuel sufficiently evaporated before the start of combustion spreads in the cylinder, and premixed combustion can be realized.

<3> Region C Since the combustion start temperature is lower than the combustion start temperature of the premixed combustion due to the decrease in the EGR gas temperature due to the decrease in the load, the intake bottom dead center is maintained to maintain the premixed combustion. The intake valve closing timing, which is also late, is advanced to approach the intake bottom dead center, and the actual compression ratio is increased to increase the combustion start temperature.

<4> Region D In the regions D and E where the load is further reduced, if the oxygen concentration is kept at the same low level as in the region C, the combustion speed is reduced, causing a problem of misfiring and a sudden increase in white smoke. Therefore, in the above-described conventional apparatus, the oxygen concentration is increased by decreasing the EGR amount, and the injection timing is advanced, so that the start of combustion is advanced to suppress misfire and unburned residue.

On the other hand, in the present invention, in the D region, the swirl is first weakened to reduce the cooling loss during the compression stroke. However, in this state, the deterioration of combustion occurs due to the decrease in the swirl ratio. Therefore, the supercharging pressure is increased by the turbocharger to increase the amount of intake air, thereby increasing the amount of working gas to reduce the swirl ratio and maintaining high angular kinetic energy even at the compression top dead center. . In addition, this increase in supercharging increases the intake air temperature and increases the actual compression ratio, thereby contributing to an increase in the combustion start temperature.

<5> Region E In an extremely low load region (for example, at the time of idling), it is necessary to further raise the combustion start temperature. Therefore, a small amount of pilot injection is performed prior to the main injection of the fuel. To increase the gas temperature in the combustion chamber at the start of main combustion. The main injection is performed after the combustion by the pilot injection is completed, and the pilot injection timing, the pilot injection amount, and the main injection timing are set so that the main combustion starts between 5 ° and 20 ° after the compression top dead center. Control. For example, in the vicinity of idle, the pilot injection timing is 35 ° before the compression top dead center, and the pilot injection amount is 1 mm ³ / s
t, the main injection timing is 3 ° before the compression top dead center.

The reason why the main injection is performed after the combustion for the pilot injection is completed is that the purpose of the pilot injection in the present invention is to increase the combustion start temperature (accordingly, the ignition of the main injection is promoted. The main injection is to be burned after the ignition delay period.

FIG. 8 shows the effect when the control is performed as shown in FIG.

In the above-described conventional apparatus, E in each of the D and E regions
NOx increases due to the decrease in the GR amount and the advance of the injection timing, and since low-temperature premixed combustion is not performed, combustion noise increases, and fuel consumption, HC, and PM (especially, SOF) also increase.

On the other hand, in the present invention, low-temperature premix combustion can be performed in each of the D and E regions, so that the NO
x, combustion noise, fuel consumption, HC, and PM (especially SOF) can all be reduced to the same level as the B region (HC slightly increases in the B region and fuel consumption improves in the B region).

FIG. 9 shows a state where the vehicle is running in the test mode.
It is a comparison of NOx emissions at equal PM emissions. In the case of low-temperature premixed combustion with a high compression ratio, N
Although the Ox emission ratio was large, the NOx emission ratio on the high load side decreased due to the reduction of the compression ratio by the conventional device. However, since the EGR rate is reduced in order to prevent a misfire or the like in a low load region, NOx on the low load side is greatly increased.

On the other hand, according to the present invention, NOx can be reduced even on the low load side. Therefore, compared with low-temperature premixed combustion at a high compression ratio, NOx emission is generally 1%.
/ 3 or less.

Next, the flowchart of FIG. 10 shows the control contents of the temperature raising control performed by activating the temperature raising control device in order to enable the premix combustion in a low load region or the like. Execute.

In step 1, the engine speed Ne, the target engine torque Torq as the engine load, and the intake manifold temperature t1 (detected by a temperature sensor, not shown) are read. In step 2, the actual oxygen concentration and a set value (for example, oxygen concentration) are read. 18%).

Here, the set value of the oxygen concentration of 18% is the upper limit of the oxygen concentration when performing the low-temperature premix combustion.
Therefore, if the actual oxygen concentration is 18% or less, it can be determined that the low-temperature premixed combustion region is in the low-temperature premixed combustion region, and if the actual oxygen concentration exceeds 18%, it is in the non-low-temperature premixed combustion region. The oxygen concentration changes by adjusting the exhaust gas recirculation rate.

Note that the actual oxygen concentration is as shown in FIG.
An air-fuel ratio sensor 38 is provided in the exhaust passage 2 and an air flow meter 39 is provided in the intake passage 3, and the values can be obtained by using the detected values of both. Needless to say, the set value (corresponding to the oxygen concentration of 18%) differs depending on the engine.

If it is in the low temperature premixed combustion region, the process proceeds to steps 3, 4, 5, and 6, where the combustion start temperature t2 and the target temperature T
1, T2, T3, T4 (where T1>T2>T3> T
The area is determined by comparing 4).

These target temperatures T1, T2, T3, T4
Is set in accordance with the engine load and the number of revolutions. For example, maps such as those shown in FIGS. It is obtained by searching based on the number Ne. For example, the operating point determined from the load and the rotational speed is U
At this time, T1 becomes approximately 855K,
T2 is approximately 845K, T3 is approximately 835K, and T4 is approximately 825K.

Here, the correspondence between the temperature range and the region shown in FIG. 7 is as follows.

Area B: t2> T1 Area C: T1 ≧ t2> T2 Area D: T2 ≧ t2> T3 Area E: T3 ≧ t2> T4 For area determination, a map as shown in FIG. However, in this case, when the intake air temperature is different from the matching temperature, the accuracy of the area determination is reduced. For example, since the combustion start temperature is lower in winter than in summer due to a change in intake air temperature, the temperature in the temperature range from T1 to T2 in summer is changed from T2 to T in winter.
If the temperature falls to 3, the control of the turbocharger and the control of the swirl control valve must be added in winter unlike in summer.

However, since the intake air temperature does not change in the map characteristics of FIG. 6, if the map characteristics of FIG. 6 are matched for summer, only the same control as in summer is performed even in winter (in this case, If so, the control of the turbocharger and the control of the swirl valve are not performed), and the combustion temperature does not reach the target temperature.

On the other hand, when the area is determined based on the combustion temperature, control according to the change in the intake air temperature can be performed (in winter, control of the turbocharger and control of the swirl control valve are added). Can be).

The combustion start temperature t2 [K] is between the intake manifold temperature (intake gas temperature) t1 [K] and t2 = t1 · εκ ^-1 where ε: compression ratio κ: specific heat ratio (≒ 1 .3) can be obtained from the intake manifold temperature t1. For example, if the intake manifold temperature is 50 ° C., then t2 = (273
The ^{+50) · 16 1.3-1 = 323 ·} 16 0.3 ≒ 850K. Of course, the intake air temperature Ta [K] and the intake negative pressure Boost
From [mmHg], the intake manifold temperature t1 may be estimated by the expression t1 = Ta · (760 + Boost) / 760.

As described above, the combustion start temperature t2 can be predicted based on the compression ratio and the intake gas temperature t1 (or the engine load affecting the intake gas temperature).

When t2> T1, the routine proceeds to steps 7 and 8, where the rotational speed Ne and the target engine torque Torq are used as shown in FIG.
1. A target EGR rate and a target main injection timing (target main injection start timing) are obtained by searching a map having the contents shown in FIG.

If T1 ≧ t2> T2, the routine proceeds to step 9, where the rotational speed Ne and the target engine torque Torq are used as shown in FIG.
After searching the map having the content 3 to obtain the target intake valve closing timing, the operations of steps 7 and 8 are executed.

The actual compression ratio is increased and the combustion start temperature is raised by advancing the intake valve closing timing normally set after the intake bottom dead center to near the bottom dead center.

Step 1 if T2 ≧ t2> T3
Proceeding to 0, 11, the rotation speed Ne and the target engine torque Tor
After searching the map containing the contents of FIGS. 14 and 15 from q to obtain the target swirl ratio and the target supercharging pressure, step 9,
Perform the operations of steps 7 and 8.

By increasing the supercharging pressure, the actual compression ratio also increases, and the combustion start temperature can be increased accordingly. Further, when the swirl ratio is reduced to weaken the swirl, the heat exchange rate of the working gas in the cylinder is reduced, the cooling loss is reduced, and the decrease in the working gas temperature is suppressed, which leads to an increase in the combustion start temperature.

Step T12 when T3 ≧ t2> T4
To find a target pilot injection amount and a target pilot injection timing by searching a map containing the contents of FIG. 16 from the rotational speed Ne and the target engine torque Torq, and then go to Step 1
The operations 0, 11, 9, 7, and 8 are executed.

The temperature of the working gas in the cylinder is increased by the pilot injection to increase the combustion start temperature.

In FIG. 16, in the small region at the lower left, data for 1 mm ³ / st as the target pilot injection amount and 35 ° before compression top dead center as the target pilot injection timing for pilot fuel injection are entered. I have.

Then, by a flow not shown, t2>
In the temperature range of T1, EGR valve control using the target EGR rate and injection timing control using the target main injection timing are performed.

On the other hand, in the temperature range T1 ≧ t2> T2, in addition to the above control, intake valve closing timing control for advancing the intake valve closing timing using the target intake valve closing timing is performed.

In the temperature range of T2 ≧ t2> T3, in addition to the above controls, swirl valve control using the target swirl ratio and supercharging pressure control using the target supercharging pressure are performed.

Further, in the temperature range of T3 ≧ t2> T4, pilot injection control using the target pilot injection amount and the target pilot injection timing is performed in addition to all the above controls.

As a mechanism capable of adjusting the intake valve closing timing (that is, a variable valve timing mechanism), for example, a mechanism as shown in FIG. 17 can be employed.

Reference numeral 60 denotes an intake valve, and 61 denotes a valve spring for biasing the intake valve 60 in the valve closing direction. A piston 63 is provided in contact with the upper end of each intake valve 60, and a piston 63 is provided by hydraulic pressure guided to a hydraulic chamber 62. 63 descends against the valve spring 61, and the intake valve 60 is opened.

The hydraulic oil discharged from the oil pump 64 is supplied from the accumulator 65 to the inlet side electromagnetic switching valves 66 and 6.
7, the oil is selectively supplied to the oil passages 68 and 69. Further, via the rotary valves 70 and 71 that rotate in synchronization with the engine rotation, # 1 cylinder, # 4 cylinder, # 2 cylinder,
Each intake valve 60 is opened sequentially by being selectively supplied to each of the three hydraulic chambers 62 of the three cylinders.

The hydraulic oil in each hydraulic chamber 62 is supplied to the oil passages 68, 6
By selectively escaping from the tank 9 through the outlet side electromagnetic switching valves 73 and 74 to the tank 75, the intake valves 60 are sequentially closed. By controlling the outlet-side electromagnetic switching valves 73 and 74, the closing timing of each intake valve 60 is freely controlled.

Therefore, in the variable valve timing mechanism 59 shown in FIG. 17, the outlet side electromagnetic switching valve 7 is controlled in accordance with the value of the target intake valve closing timing obtained by executing the flow of FIG.
3 and 74 may be controlled.

As described above, in the present invention, the low-temperature premixed combustion region can be expanded to the high load side by setting the compression ratio to 16 or less, which is lower than that of the conventional device. Further, when the combustion start temperature falls within a temperature range in which the low-temperature premix combustion can be performed at a target temperature T1 or lower, the main combustion is started at a predetermined time (for example, between 5 ° and 20 ° after the top dead center). In addition, the main injection timing is controlled and the temperature rise control according to the temperature range is performed so that the combustion start temperature exceeds the target temperature T1, so that the low-temperature premixed combustion can be performed up to the low load range (low temperature range). That is, the low-temperature premixed combustion region can be expanded. As a result, NOx, combustion noise, fuel consumption, H
Both C and PM were improved over the above-mentioned conventional device (see FIG. 8).

Next, another embodiment of the present invention will be described with reference to the flowchart of FIG.

This flowchart is described corresponding to FIG. 10 which is the flowchart of the first embodiment.
In the figure, the same steps as those in FIG. 10 are denoted by the same step numbers.

This embodiment is directed to an engine having a plurality of, for example, two exhaust valves per cylinder. In order to raise the combustion start temperature t2 even at a very low temperature, one of the plurality of exhaust valves is opened during the intake stroke and the EGR valve is closed, so that the high-temperature exhaust gas flows directly back into the combustion chamber. , Raise the gas temperature in the combustion chamber,
Even when the combustion start temperature t2 is lower than the target temperature T4, low-temperature premix combustion can be realized.

For example, as shown in FIG.
It is assumed that the R rate is 100% (the intake amount and the EGR gas amount are the same). In this case, even if the cooling water temperature decreases, the maximum compression temperature (=
In order to maintain the combustion start temperature t2) at a predetermined value, it is necessary to increase the pre-compression temperature (= intake manifold temperature t1) as the cooling water temperature decreases as shown in the figure.

Here, the pre-compression temperature flows back into the combustion chamber by opening one of the exhaust valves during the intake stroke and the low-temperature exhaust gas flowing through the EGR valve and the low-temperature air flowing through the intake passage. It is determined by the balance with the hot gas. Therefore, in order to make the pre-compression temperature the characteristic shown in the drawing, the flow rate of one exhaust valve to be opened during the intake stroke should be increased and the intake valve flow rate and the EGR valve flow rate should be decreased as the cooling water temperature decreases.

Returning to the flowchart of FIG. 18, FIG.
Only the step 21 is different. That is, when the combustion start temperature t2 is equal to or lower than the target temperature T4, the process proceeds from step 6 to step 21, in which a table set in the same manner as in FIG. 19 is searched based on the cooling water temperature, and the target EGR valve flow rate, target exhaust valve The flow rate and the target intake valve flow rate are obtained. Thereafter, as in the first embodiment, Step 1
The operations of 2, 10, 11, 9, 7, and 8 are executed.

Then, the target flow rates obtained as described above are used as EGR values.
In a flow (not shown), the EGR valve opening and the intake valve opening are controlled so as to flow through the valve, the exhaust valve, and the intake valve, and the exhaust valve opening during the intake stroke is controlled.

Further, since the EGR gas flows back into the combustion chamber by opening one of the exhaust valves during the intake stroke, the swirl in the combustion chamber is weakened. An exhaust port on the valve side is formed as a helical port so that the exhaust gas flowing backward generates a swirling motion in the combustion chamber.

As described above, when the combustion start temperature t2 becomes equal to or lower than the target temperature T4, the low-temperature EGR gas and the low-temperature intake air flowing into the combustion chamber via the EGR valve are reduced, and one of the two is reduced during the intake stroke. Opening the exhaust valve increases the amount of hot EGR gas that flows back into the combustion chamber. Moreover, at this time, the exhaust port of the exhaust valve which is opened in the intake stroke is formed as a helical port in order to prevent the swirl in the combustion chamber from lowering due to the opening of the exhaust valve. As a result, even at a very low temperature where the combustion start temperature t2 is equal to or lower than the target temperature T4,
Low-temperature premixed combustion can be realized, thereby further improving exhaust performance at low water temperature.

Although the respective flow rates of the EGR valve, one of the exhaust valves and the intake valve are changed in accordance with the cooling water temperature, the simplest control is to set one of the exhaust valves to a predetermined opening degree during the intake stroke. , And the EGR valve and the intake valve may be closed to the minimum opening (this corresponds to the characteristic at the left end in FIG. 19).

In the embodiment, the case where the region is determined based on the combustion temperature has been described. However, the region may be determined based on the rotation speed and the load as described above with reference to FIG. FIGS. 24 and 25 show flowcharts in this case. Note that FIG.
4 and FIG. 25 are described corresponding to FIG. 10 and FIG. 18 which are flowcharts of the first and second embodiments.
The same steps as those in FIG. 18 are denoted by the same step numbers.

To explain this, in FIG. 6, the load defining the boundary between the region B and the region C is represented by the first target load,
With the load that determines the boundary of the region as the second target load, and the load that determines the boundary between the region D and the region E as the third target load, the temperature rise control is performed as follows.

(1) When the load is equal to or less than the first target load and
A load range lower than the target load and higher than the second target load (C
(Region 33), the temperature increase control is performed by advancing the closing timing of the intake valve (steps 33 and 9).

(2) When the load is equal to or less than the second target load and
A load range higher than the third target load, which is lower than the target load (D
(Region), the intake valve closing timing is advanced, the swirl ratio is reduced, and the supercharging pressure is further increased to perform the temperature increase control (steps 34, 10, 11, 9).

(3) When the load is equal to or less than the third target load and
A load range lower than the target load and higher than the fourth target load (E
Region), the intake valve closing timing is advanced, the swirl ratio is reduced, the supercharging pressure is increased, and pilot injection is performed prior to the main fuel injection, and combustion by the pilot injection is performed before the main injection. By setting the pilot injection timing so as to end the process, the temperature increase control is performed (step 3
5, 12, 10, 11, 9), the main fuel injection timing is controlled so that the main injected fuel is ignited after the end of the main fuel injection.

In the above embodiment, the oxygen amount is changed by the EGR rate. However, the present invention is not limited to this case. For example, the oxygen amount can be changed by using an oxygen permeable membrane. It is.

[Brief description of the drawings]

FIG. 1 is a control system diagram of a first embodiment.

FIG. 2 is a characteristic diagram showing characteristics of a combustion start temperature according to an engine load.

FIG. 3 is a characteristic diagram for explaining a combustion start temperature region and a target attainment temperature of low-temperature premixed combustion.

FIG. 4 is a characteristic diagram of temperature rise when combustion start timings are different.

FIG. 5 is a characteristic diagram of a heat release rate pattern when combustion start times are different.

FIG. 6 is a region diagram in which a control region for an engine speed and an engine load is divided.

FIG. 7 is a characteristic diagram for explaining control for each area in FIG. 6;

8 is a characteristic diagram showing emission characteristics of NOx and the like when the control in FIG. 7 is performed.

FIG. 9 is a characteristic diagram comparing NOx emission amounts at the same PM during mode driving.

FIG. 10 is a flowchart for explaining control contents.

FIG. 11 is a characteristic diagram of a target EGR rate.

FIG. 12 is a characteristic diagram of a target main injection timing.

FIG. 13 is a characteristic diagram of a target intake valve closing timing.

FIG. 14 is a characteristic diagram of a target swirl ratio.

FIG. 15 is a characteristic diagram of a target supercharging pressure.

FIG. 16 is a characteristic diagram of a target pilot injection amount and a target pilot injection timing.

FIG. 17 is a schematic configuration diagram of a variable valve timing mechanism.

FIG. 18 is a flowchart for explaining control contents according to the second embodiment.

FIG. 19 is a characteristic diagram of a target EGR valve flow rate, a target exhaust valve flow rate, and a target intake valve flow rate.

FIG. 20 is a map in which a first target temperature is set.

FIG. 21 is a map in which a second target temperature is set.

FIG. 22 is a map in which a third target temperature is set.

FIG. 23 is a map in which a fourth target temperature is set.

FIG. 24 is a flowchart for explaining control contents of the third embodiment.

FIG. 25 is a diagram corresponding to a claim of the first invention.

FIG. 26 is a diagram corresponding to the claims of the fifteenth invention.

[Description of Signs] 6 EGR valve 17 Nozzle 33 Accelerator opening sensor 34 Crank angle sensor 41 Control unit 52 Turbocharger

──────────────────────────────────────────────────続 き Continued on the front page (51) Int.Cl. ⁶ Identification code FI F02D 41/02 370 F02D 41/02 370 41/38 41/38 B 41/40 41/40 E 43/00 301 43/00 301J 301N 301U 301Z 301R F02M 25/07 570 F02M 25/07 570J

Claims (20)

[Claims] 1. A diesel engine for performing low-temperature premixed combustion at a low compression ratio, a fuel injection valve having a variable fuel injection timing, a temperature increase control device for increasing a temperature of a working gas in a combustion chamber, and combustion start. Means for predicting the atmospheric temperature in the cylinder at the time of combustion, means for determining whether the atmospheric temperature in the cylinder at the start of combustion is in a region lower than a first target temperature for maintaining low-temperature premixed combustion, When it is determined that the ambient temperature in the cylinder is lower than the first target temperature, the temperature raising control device is operated so that the ambient temperature in the cylinder at the start of combustion exceeds the first target temperature. Means for adjusting the fuel injection timing of the fuel injector so that the rate of increase of the combustion temperature is equal to or higher than a predetermined value. 2. The diesel engine combustion control device according to claim 1, wherein an ambient temperature in the cylinder at the start of the combustion is predicted based on a compression ratio and an engine load. 3. The combustion control apparatus for a diesel engine according to claim 1, wherein the atmospheric temperature in the cylinder at the start of the combustion is predicted based on a compression ratio and an intake gas temperature. 4. In order for the temperature rise rate of the main combustion to be equal to or higher than the predetermined value, the combustion is started 5 to 20 after compression top dead center.
The combustion control device for a diesel engine according to any one of claims 1 to 3, wherein the fuel injection timing is controlled so as to fall within the range of ゜. 5. An apparatus according to claim 1, further comprising an intake valve closing timing control mechanism as said temperature raising control device, wherein a second target temperature of said cylinder at the start of combustion is equal to or lower than said first target temperature and lower than said first target temperature. The combustion control device for a diesel engine according to any one of claims 1 to 4, wherein when the temperature is higher than the temperature, the temperature increase control is performed by advancing a closing timing of the intake valve. . 6. A swirl control mechanism is further provided as the temperature raising control device, wherein an ambient temperature in the cylinder at the start of the combustion is lower than the second target temperature and lower than a third target temperature lower than the second target temperature. The combustion control device for a diesel engine according to claim 5, wherein when the temperature is also high, the temperature increase control is performed by advancing the intake valve closing timing and lowering a swirl ratio. 7. A third target, further comprising a supercharging pressure control mechanism as the temperature raising control device, wherein an ambient temperature in the cylinder at the start of the combustion is equal to or lower than the second target temperature and lower than the second target temperature. The diesel engine according to claim 5, wherein when the temperature is higher than the temperature, the temperature increase control is performed by advancing the intake valve closing timing, lowering a swirl ratio, and further increasing a supercharging pressure. Engine combustion control device. 8. A fourth target temperature, further comprising a fuel injection control mechanism as said temperature raising control device, wherein an in-cylinder ambient temperature at the start of combustion is equal to or lower than said third target temperature and lower than said third target temperature. When it is higher than the above, the intake valve closing timing is advanced, the swirl ratio is lowered, the pilot injection is performed prior to the main injection of the fuel, and the pilot injection is performed so that the combustion by the pilot injection ends before the main injection. The combustion control device for a diesel engine according to claim 6, wherein the temperature increase control is performed by setting an injection timing. 9. A fourth target temperature, wherein a fuel injection control mechanism is further provided as the temperature increase control device, and an atmosphere temperature in the cylinder at the start of the combustion is equal to or lower than the third target temperature and lower than the third target temperature. When it is higher than the above, the intake valve closing timing is advanced, the swirl ratio is lowered, the supercharging pressure is increased, and the pilot injection is performed prior to the main injection of the fuel. The diesel engine combustion control device according to claim 6 or 7, wherein the temperature increase control is performed by setting a pilot injection timing so as to end before. 10. The diesel engine combustion control device according to claim 8, wherein the main fuel injection timing is controlled so that the main fuel is ignited after the main fuel injection is completed. 11. A valve opening / closing mechanism capable of arbitrarily controlling the opening / closing timing of one of a plurality of exhaust valves provided for one cylinder as the temperature rise control device, and an amount of exhaust gas recirculated into intake air. An EGR valve for controlling the intake valve closing timing when the in-cylinder ambient temperature at the start of combustion is equal to or lower than the fourth target temperature, lowering the swirl ratio, increasing the supercharging pressure, Performing the pilot injection so that the main injection is performed after the combustion of the pilot injection is completed, and performing the temperature increase control by opening the one exhaust valve and closing the EGR valve in an intake stroke. The combustion control device for a diesel engine according to any one of claims 8 to 10. 12. An exhaust valve opening / closing mechanism capable of arbitrarily controlling the opening / closing timing of one of a plurality of exhaust valves provided for one cylinder as the temperature rise control device, and optionally setting the opening / closing timing of an intake valve. An intake valve control mechanism that can be adjusted to a predetermined value, and an EGR valve that controls the amount of exhaust gas recirculated during intake. When the cylinder ambient temperature at the start of combustion is equal to or lower than the fourth target temperature, the intake valve closing timing To lower the swirl ratio,
The supercharging pressure is increased, the pilot injection is performed so that the main injection is performed after the combustion of the pilot injection is completed, and the amounts of the EGR gas and intake air flowing into the combustion chamber via the EGR valve are changed to the cooling water temperature. The temperature increase control is performed by decreasing one as the cooling water temperature decreases and increasing the amount of EGR gas flowing back to the combustion chamber by opening one exhaust valve in the intake stroke as the cooling water temperature decreases. The combustion control device for a diesel engine according to any one of 8 to 10. 13. The diesel engine combustion control device according to claim 11, wherein an exhaust port on the one exhaust valve side opened in the intake stroke is formed as a helical port. 14. The diesel engine combustion control according to claim 1, wherein the target temperature is set according to the load so that the temperature decreases as the engine load decreases. apparatus. 15. A diesel engine for performing low-temperature premixed combustion at a low compression ratio, a fuel injection valve whose fuel injection timing is variable, a temperature increase control device for increasing a temperature of a working gas in a combustion chamber, and combustion start. Means for detecting an engine load that influences the cylinder internal temperature at the time of the engine load, wherein the engine load is lower than or equal to a first target load for maintaining low-temperature premixed combustion and lower than the first target load. Means for determining whether the engine load is in a high load range, and when the engine load is determined to be in a load range lower than the first target load and higher than the second target load, the temperature raising control device is operated. And means for adjusting the fuel injection timing of the fuel injection valve such that the rate of increase of the combustion temperature is equal to or higher than a predetermined value. Control device. 16. A swirl control mechanism is further provided as the temperature raising control device, and when the load range is lower than the second target load and higher than a third target load lower than the second target load, the intake valve is controlled. The combustion control device for a diesel engine according to claim 15, wherein the temperature increase control is performed by advancing a closing timing and lowering a swirl ratio. 17. The apparatus according to claim 17, further comprising a supercharging pressure control mechanism as said temperature raising control device, wherein said supercharging pressure control device is provided at a pressure equal to or less than said second target load and
When the load range is higher than the third target load that is lower than the target load, the intake valve closing timing is advanced, and the swirl ratio is lowered,
The combustion control device for a diesel engine according to claim 15, wherein the temperature increase control is performed by further increasing a supercharging pressure. 18. A fuel injection control mechanism is further provided as the temperature increase control device, and when the load range is lower than the third target load and higher than a fourth target load lower than the third target load, the intake air is controlled. By advancing the valve closing timing, lowering the swirl ratio, further performing pilot injection before the main injection of fuel, and setting the pilot injection timing so that combustion by pilot injection ends before the main injection And the temperature raising control is performed.
7. The combustion control device for a diesel engine according to 6. 19. A fuel injection control mechanism is further provided as the temperature increase control device, and when the load range is lower than the third target load and higher than a fourth target load lower than the third target load, the intake air is controlled. Pilot the valve closing timing, lower the swirl ratio, increase the supercharging pressure, perform pilot injection before the main fuel injection, and terminate the pilot injection combustion before the main injection. 17. The diesel engine combustion control device according to claim 15, wherein the temperature increase control is performed by setting an injection timing. 20. The combustion control system for a diesel engine according to claim 18, wherein the main fuel injection timing is controlled so that the main fuel is ignited after the main fuel injection is completed.