JP3094992B2 - Internal combustion engine - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

[0001] The present invention relates to an internal combustion engine.

[0002]

Conventionally than internal combustion engines, for example exhaust gas recirculation and engine exhaust passage and the engine intake passage in order to suppress the generation of the NO _x in the diesel engine (hereinafter, referred to as EGR) connected by passages, the Exhaust gas, that is, EGR gas, is recirculated through the EGR passage into the engine intake passage. In this case, the EGR gas has a relatively high specific heat, and therefore can absorb a large amount of heat.
The combustion temperature in the combustion chamber decreases as the GR gas amount increases, that is, as the EGR rate (EGR gas amount / (EGR gas amount + intake air amount)) increases. When the combustion temperature is lowered to decrease the generated amount of NO _x, thus the generation amount of the more NO _x to be increased EGR rate is lowered.

[0003] It has been found that can reduce the generation amount of the NO _x Thus conventionally increasing the EGR rate. However, when the EGR rate is increased, the soot generation amount, that is, smoke, starts to increase rapidly when the EGR rate exceeds a certain limit. In this regard, it has conventionally been considered that if the EGR rate is further increased, the smoke will increase indefinitely. Therefore, the smoke starts to increase rapidly.
The GR rate is considered to be the maximum allowable limit of the EGR rate.

Therefore, conventionally, the EGR rate is set within a range not exceeding the maximum allowable limit. The maximum allowable EGR rate varies considerably depending on the type of engine and fuel, but is approximately 30 to 50%.
Therefore, in a conventional diesel engine, the EGR rate is at most 3
It is reduced from 0% to about 50%.

As described above, conventionally, it has been considered that the maximum allowable limit exists for the EGR rate.
If the R rate is within the range not exceeding this maximum allowable limit, NO
_It was set so that the amount of _x and smoke generated was as small as possible. However, in this way EGR
Rate that there is a limit to the reduction of the NO _x and the amount of generated NO _x and the amount of smoke produced also defined to be as small as possible of smoke, in fact still a significant amount of N
At present, O _x and smoke are generated.

However, if the EGR rate is made larger than the maximum allowable limit in the course of research on the combustion of a diesel engine, the smoke rapidly increases as described above. However, the amount of generated smoke has a peak, and the peak exceeds this peak. EGR
When the rate is further increased, the smoke starts to decrease rapidly, and when the EGR rate is increased to 70% or more during idling operation, and when the EGR gas is cooled strongly, the smoke is reduced to about 55% or more. It was found that it was almost zero, that is, almost no soot was generated. In this case, N
Generation amount of O _x is also found that a very small amount.
After that, the reason why no soot was generated was examined based on this finding, and as a result, unprecedented soot and NO
_This has led to the construction of a new combustion system capable of simultaneously reducing _x . This new combustion system will be described in detail later, but in short, it is basically based on stopping the growth of hydrocarbons in the middle stage until the hydrocarbons grow into soot.

That is, as a result of repeated experimental studies, it has been found that when the temperature of the fuel and the surrounding gas during combustion in the combustion chamber is lower than a certain temperature, the growth of hydrocarbons is stopped at a halfway stage before reaching soot. However, when the temperature of the fuel and the gas around it rises above a certain temperature, the hydrocarbons grow into soot at a stretch. In this case, the temperature of the fuel and the surrounding gas is greatly affected by the heat absorbing action of the gas around the fuel when the fuel is burned, and the amount of heat absorbed by the gas around the fuel is adjusted according to the calorific value at the time of burning the fuel. As a result, the temperature of the fuel and the surrounding gas can be controlled.

Accordingly, if the temperature of the fuel and the surrounding gas during combustion in the combustion chamber is suppressed to a temperature below the temperature at which the growth of hydrocarbons stops halfway, soot will not be generated, and the fuel during combustion in the combustion chamber and its surroundings will not be generated. Can be suppressed to a temperature below the temperature at which the growth of hydrocarbons stops halfway, by adjusting the amount of heat absorbed by the gas around the fuel. On the other hand, hydrocarbons whose growth has stopped halfway before reaching soot can be easily purified by post-treatment using an oxidation catalyst or the like. This is the basic idea of a new combustion system. An internal combustion engine employing this new combustion system has already been filed by the present applicant (Japanese Patent Application No. 9-305850).

[0009]

However, this new combustion system requires that the EGR rate be approximately 55% or more. However, the EGR rate can be increased to about 55% or more when the intake air amount is relatively small, that is, when the engine load is relatively low. When the intake air amount exceeds a certain amount, this new combustion is no longer performed. Can not do. Therefore, when the amount of intake air exceeds a certain amount, it is necessary to switch to the conventional combustion.

However, the form of combustion is different between the new combustion and the conventional combustion. Therefore, the optimum fuel injection start timing under the new combustion and the optimal fuel injection under the conventional combustion are different. Is completely different from the fuel injection start timing. In other words, new combustion requires a large amount of E
Since the combustion is performed in the presence of the GR gas, the combustion becomes slow. Therefore, if the fuel is started to be injected at the optimal injection start timing under the conventional combustion, the injection start timing becomes too late, so that good combustion is not performed. A misfire will result. Therefore, when performing new combustion, it is necessary to make the fuel injection start timing earlier than when performing conventionally performed combustion.

Further, when switching between new combustion and conventional combustion, the EGR rate in the combustion chamber must be instantaneously switched, but it is actually difficult to instantaneously switch the EGR rate in the combustion chamber. In fact, the EGR rate in the combustion chamber gradually changes. That is, when the combustion is switched from the new combustion to the conventional combustion, the combustion mode gradually changes from the new combustion mode to the conventional combustion mode. In this case, the combustion mode is gradually changed from the conventional combustion mode to a new combustion mode.

Therefore, when switching from the new combustion to the conventional combustion, the fuel injection start timing and the fuel injection amount are changed from the injection start timing and the injection amount suitable for the new combustion to the conventional combustion. It is preferable to gradually switch the injection start timing and the injection amount suitable for the fuel injection, and when the conventional combustion is switched to the new combustion, the fuel injection start timing and the fuel injection amount are conventionally performed. It can be said that it is preferable to gradually switch from the injection start timing and injection amount suitable for combustion to the injection start timing and injection amount suitable for new combustion.

An object of the present invention is to provide an internal combustion engine that performs optimal injection control when switching between new combustion and conventional combustion.

[0014]

In order to achieve the above object, according to a first aspect of the present invention, as the amount of inert gas in the combustion chamber increases, the amount of soot generated gradually increases and reaches a peak. When the amount of the inert gas is further increased, the temperature of the fuel during combustion in the combustion chamber and the temperature of the surrounding gas become lower than the temperature at which soot is generated, and soot is hardly generated. A first combustion in which a valve is disposed, and the amount of inert gas in the combustion chamber is larger than the amount of inert gas at which the amount of generated soot becomes a peak, and soot is hardly generated;
Switching means for selectively switching between the second combustion in which the amount of inert gas in the combustion chamber is smaller than the amount of inert gas in which the amount of soot generation reaches a peak, and switching from the first combustion to the second combustion; When it is, the fuel injection start timing is delayed, and when it is switched from the second combustion to the first combustion, the fuel injection start timing is advanced.

In the second invention, in the first invention,
At least at the time of switching between the first combustion and the second combustion, the second combustion is performed by performing the pilot injection prior to the main injection, and when the first combustion is switched to the second combustion, the first combustion is performed. In addition to delaying the injection start timing of the main injection as compared to the injection start timing when the combustion of
When the combustion is switched from the first combustion to the first combustion, the injection start timing is advanced earlier than the injection start timing of the main injection.

In the third invention, in the second invention,
When the first combustion is switched to the second combustion, the injection amount of the main injection is gradually increased and the injection amount of the pilot injection is gradually decreased. In a fourth aspect based on the third aspect, when the first combustion is switched to the second combustion, the injection start timing of the pilot injection is changed from the injection start timing in the first combustion to the operation of the engine during the second combustion. It is gradually changed toward a predetermined pilot injection start timing according to the state.

According to a fifth aspect, in the third aspect,
When the mode is switched from the first combustion to the second combustion, the injection start timing of the main injection is gradually advanced. In a sixth aspect based on the second aspect, when the mode is switched from the second combustion to the first combustion, the injection amount of the main injection is gradually reduced and the injection amount of the pilot injection is gradually increased. .

In the seventh invention, in the sixth invention,
When switching from the second combustion to the first combustion, the injection start timing of the pilot injection is gradually changed toward the injection start timing in the first combustion. In an eighth aspect based on the sixth aspect, the injection start timing of the main injection is gradually delayed when the mode is switched from the second combustion to the first combustion.

According to a ninth aspect, in the first aspect,
A recirculation device is provided for recirculating exhaust gas discharged from the combustion chamber into the engine intake passage, and the inert gas comprises the recirculated exhaust gas. In a tenth aspect based on the ninth aspect, the exhaust gas recirculation rate in the first combustion state is approximately 5%.
5% or more.

According to an eleventh aspect, in the first aspect, a catalyst having an oxidizing function is disposed in the engine exhaust passage. In 11 th invention is a 12 th invention, the catalyst consists of one at least of the oxidation catalyst, three-way catalyst or _the NO _x absorbent. In a thirteenth aspect based on the first aspect, the operating range of the engine is divided into a first operating range on the low load side and a second operating range on the high load side, and the first combustion is performed in the first operating range. In the second operation region, the second combustion is performed.

[0021]

FIG. 1 shows a case where the present invention is applied to a four-stroke compression ignition type internal combustion engine. Referring to FIG. 1, 1 is an engine body, 2 is a cylinder block, 3 is a cylinder head, 4 is a piston, 5 is a combustion chamber, 6 is an electrically controlled fuel injection valve, 7 is an intake valve, 8 is an intake port, 9 Denotes an exhaust valve, and 10 denotes an exhaust port. The intake port 8 is connected to a surge tank 12 via a corresponding intake branch pipe 11. Be linked. An inlet of the compressor 16 is connected to an air cleaner 18 via an air suction pipe 17, and a throttle valve 20 driven by a step motor 19 is arranged in the air suction pipe 17.

On the other hand, the exhaust port 10 is connected to the exhaust manifold 2.
1 and the exhaust turbocharger 15 via the exhaust pipe 22
The exhaust gas turbine 23 is connected to a catalytic converter 26 having a built-in catalyst 25 having an oxidizing function via an exhaust pipe 24. An air-fuel ratio sensor 27 is arranged in the exhaust manifold 21.

The exhaust pipe 28 connected to the outlet of the catalytic converter 26 and the air suction pipe 17 downstream of the throttle valve 20 are connected to each other via an exhaust gas recirculation (hereinafter referred to as EGR) passage 29, and are connected to an EGR passage. An EGR control valve 31 driven by a step motor 30 is arranged in the inside 29. In the EGR passage 29, an intercooler 32 for cooling the EGR gas flowing in the EGR passage 29 is arranged. In the embodiment shown in FIG. 1, the engine cooling water is guided into the intercooler 32, and the engine cooling water
The GR gas is cooled.

On the other hand, the fuel injection valve 6 is connected to a fuel reservoir, a so-called common rail 34, via a fuel supply pipe 33. Fuel is supplied into the common rail 34 from an electric control type variable discharge fuel pump 35, and the fuel supplied into the common rail 34 is supplied to the fuel injection valve 6 through each fuel supply pipe 33. A fuel pressure sensor 36 for detecting the fuel pressure in the common rail 34 is attached to the common rail 34, and the fuel pump 35 is controlled so that the fuel pressure in the common rail 34 becomes the target fuel pressure based on the output signal of the fuel pressure sensor 36. Is controlled.

The electronic control unit 40 is composed of a digital computer, and is connected to a ROM (Read Only Memory) 42, a RAM (Random Access Memory) 43, a CPU (Microprocessor) 44, an input port 45, An output port 46 is provided. The output signal of the air-fuel ratio sensor 27 is input to the input port 45 via the corresponding AD converter 47, and the output signal of the fuel pressure sensor 36 is also input to the input port 45 via the corresponding AD converter 47. A load sensor 51 that generates an output voltage proportional to the depression amount L of the accelerator pedal 50 is connected to the accelerator pedal 50, and the output voltage of the load sensor 51 is input to the input port 4 via the corresponding AD converter 47.
5 is input. The input port 45 is connected to a crank angle sensor 52 that generates an output pulse every time the crankshaft rotates, for example, by 30 °. On the other hand, the output port 46 is connected to the fuel injection valve 6,
It is connected to a throttle valve control step motor 19, an EGR control valve control step motor 30 and a fuel pump 35.

FIG. 2 shows the throttle valve 2 when the engine is under low load operation.
Change in the output torque when changing the air-fuel ratio A / F (abscissa in FIG. 2) by changing the opening and the EGR rate of 0, and smoke, HC, CO, a change in emission of the NO _x It shows the experimental example shown. As can be seen from FIG. 2, in this experimental example, the smaller the air-fuel ratio A / F, the higher the EGR rate. When the air-fuel ratio A / F is smaller than the stoichiometric air-fuel ratio (≒ 14.6), the EGR rate is 65% or more.

As shown in FIG. 2, when the air-fuel ratio A / F is reduced by increasing the EGR rate, the smoke is reduced when the EGR rate becomes about 40% and the air-fuel ratio A / F becomes about 30. The generation starts to increase. Next, when the EGR rate is further increased and the air-fuel ratio A / F is reduced, the amount of smoke generated sharply increases and reaches a peak. Next, when the EGR rate is further increased and the air-fuel ratio A / F is reduced, the smoke is sharply reduced. When the EGR rate is increased to 65% or more and the air-fuel ratio A / F is around 15.0, the smoke becomes almost zero. . That is, almost no soot is generated. At this time, the output torque of the engine slightly decreases, and N
The generation amount of O _x is considerably reduced. On the other hand, at this time, HC,
The amount of generated CO starts to increase.

FIG. 3A shows a change in the combustion pressure in the combustion chamber 5 when the amount of generated smoke is the largest when the air-fuel ratio A / F is around 18, and FIG. 3B shows the air-fuel ratio A / F. The graph shows a change in the combustion pressure in the combustion chamber 5 when F is around 13 and the amount of generated smoke is almost zero. As can be seen by comparing FIG. 3 (A) and FIG. 3 (B), in the case of FIG. 3 (B) where the amount of smoke generation is almost zero, FIG.
It can be seen that the combustion pressure is lower than in the case shown in (A).

The following can be said from the experimental results shown in FIGS. That is, first, the air-fuel ratio A / F is 1
FIG. 2 when the smoke generation amount is almost zero at 5.0 or less.
As shown in ₍₂₎ , the generation amount of NOx is considerably reduced. N
That the generation amount of O _x produced falls means that the combustion temperature in the combustion chamber 5 is reduced, thus the combustion temperature in the combustion chamber 5 becomes low when the soot is hardly generated I can say. The same can be said from FIG. That is, in the state shown in FIG. 3B where almost no soot is generated, the combustion pressure is low.
The combustion temperature inside is low.

Second, when the amount of generated smoke, that is, the amount of generated soot becomes almost zero, as shown in FIG.
Emissions increase. This means that hydrocarbons are emitted without growing to soot. That is, the linear hydrocarbons and aromatic hydrocarbons contained in the fuel as shown in FIG. 4 are thermally decomposed when the temperature is increased in a state of lack of oxygen, soot precursors are formed, and then mainly, Soot consisting of a solid aggregate of carbon atoms is produced. In this case, the actual soot production process is complicated, and it is not clear what form the soot precursor takes, but in any case, the hydrocarbon as shown in FIG. It will grow to soot. Therefore, as described above, when the amount of generated soot becomes substantially zero, the emission amounts of HC and CO increase as shown in FIG. 2, but HC at this time is a precursor of soot or a hydrocarbon in a state before it. .

When these considerations based on the experimental results shown in FIGS. 2 and 3 are summarized, when the combustion temperature in the combustion chamber 5 is low, the amount of soot generation becomes almost zero. Is discharged from the combustion chamber 5. As a result of further detailed experimental study on this, if the temperature of the fuel and the surrounding gas in the combustion chamber 5 is lower than a certain temperature, the growth process of the soot is stopped halfway, that is, the soot is It was found that no soot was generated, and soot was generated when the temperature of the fuel and its surroundings in the combustion chamber 5 exceeded a certain temperature.

The temperature of the fuel and its surroundings when the process of producing hydrocarbons is stopped in the state of the soot precursor, that is, the above-mentioned certain temperature, depends on various factors such as the type of fuel, air-fuel ratio and compression ratio. Although the change can not be said that how many times since this certain temperature has a generation amount and the closely related of the nO _x, therefore this certain temperature is defined to a certain degree from the generation amount of the nO _x be able to. That is, the fuel and the gas temperature surrounding it at the time of combustion and the greater the EGR rate, decreases, the amount of the NO _x is reduced. Generation amount at this time NO _x is soot is hardly generated when it is around or less 10 ppm. Therefore, the above certain temperature is NO
It almost coincides with the temperature when the amount of generated _x is about 10 p.pm or less.

Once soot has been produced, it cannot be purified by post-treatment using a catalyst having an oxidizing function. On the other hand, the soot precursor or the hydrocarbon in the state before the soot can be easily purified by a post-treatment using a catalyst having an oxidation function. Considering the post-treatment with a catalyst having an oxidation function as described above, it is extremely difficult to discharge hydrocarbons from the combustion chamber 5 in the state of a precursor of soot or in the state before the soot or in the form of soot from the combustion chamber 5. There is a big difference. The new combustion system employed in the present invention discharges hydrocarbons from the combustion chamber 5 in the form of a soot precursor or previous state without producing soot in the combustion chamber 5 and removes the hydrocarbons. The core is to oxidize with a catalyst having an oxidation function.

Now, in order to stop the growth of hydrocarbons before the soot is generated, the temperature of the fuel and the surrounding gas during combustion in the combustion chamber 5 is set to a temperature lower than the temperature at which the soot is generated. It needs to be suppressed. In this case, it has been found that the endothermic effect of the gas around the fuel when the fuel burns has an extremely large effect on suppressing the temperature of the fuel and the gas around the fuel.

That is, if there is only air around the fuel, the evaporated fuel immediately reacts with oxygen in the air and burns. In this case, the temperature of the air separated from the fuel does not rise so much, and only the temperature around the fuel becomes extremely high locally. That is, at this time, the air separated from the fuel hardly absorbs the heat of combustion heat of the fuel. In this case, since the combustion temperature locally becomes extremely high, the unburned hydrocarbons that have received the combustion heat will generate soot.

On the other hand, when fuel is present in a mixed gas of a large amount of inert gas and a small amount of air, the situation is slightly different.
In this case, the fuel vapor diffuses to the surroundings, reacts with oxygen mixed in the inert gas, and burns. In this case, the combustion temperature is not increased so much because the combustion heat is absorbed by the surrounding inert gas. That is, the combustion temperature can be kept low. That is, the presence of the inert gas plays an important role in suppressing the combustion temperature, and the combustion temperature can be kept low by the endothermic effect of the inert gas.

In this case, in order to suppress the temperature of the fuel and the surrounding gas to a temperature lower than the temperature at which the soot is generated, an amount of the inert gas that can absorb a sufficient amount of heat to do so is required. . Therefore, if the fuel amount increases, the required amount of inert gas increases accordingly. In this case, as the specific heat of the inert gas increases, the endothermic effect becomes stronger. Therefore, the inert gas preferably has a higher specific heat. In this regard, it can be said that it is preferable to use EGR gas as the inert gas since CO ₂ and EGR gas have relatively large specific heats.

FIG. 5 shows the relationship between the EGR rate and smoke when EGR gas is used as the inert gas and the degree of cooling of the EGR gas is changed. That is, in FIG. 5, a curve A indicates that the EGR gas temperature is substantially 9
Curve B shows the case where the EGR gas is cooled by a small cooling device, and curve C shows the case where the temperature is maintained at 0 ° C.
Indicates a case where the EGR gas is not forcibly cooled.

As shown by the curve A in FIG. 5, when the EGR gas is cooled strongly, the amount of soot generation peaks at a point where the EGR rate is slightly lower than 50%. Above a percentage, little soot is generated. On the other hand, as shown by the curve B in FIG. 5, when the EGR gas is slightly cooled, the soot generation amount reaches a peak at a point where the EGR rate is slightly higher than 50%. In this case, the EGR rate is increased to about 65% or more. If so, almost no soot is generated.

As shown by the curve C in FIG.
When the R gas is not forcibly cooled, the EGR rate becomes 5
The soot generation amount peaks near 5%, and in this case, if the EGR rate is set to approximately 70% or more, soot is hardly generated. FIG. 5 shows the amount of smoke generated when the engine load is relatively high. When the engine load decreases, the EGR rate at which the amount of soot peaks slightly decreases, and the EGR rate at which soot is hardly generated is reduced. Also lowers slightly. As described above, the lower limit of the EGR rate at which almost no soot is generated varies depending on the degree of cooling of the EGR gas and the engine load.

FIG. 6 shows a mixture of EGR gas and air necessary to make the temperature of the fuel and the surrounding gas during combustion lower than the temperature at which soot is generated when EGR gas is used as the inert gas. It shows the gas amount, the ratio of air in the mixed gas amount, and the ratio of EGR gas in the mixed gas. In FIG. 6, the vertical axis indicates the total intake gas amount sucked into the combustion chamber 5, and the dashed line Y indicates the total intake gas amount that can be sucked into the combustion chamber 5 when supercharging is not performed. ing. The horizontal axis indicates the required load.

Referring to FIG. 6, the proportion of air, that is, the amount of air in the mixed gas, indicates the amount of air necessary to completely burn the injected fuel. That is, in the case shown in FIG. 6, the ratio between the air amount and the injected fuel amount is the stoichiometric air-fuel ratio. On the other hand, in FIG. 6, the ratio of the EGR gas, that is, the amount of the EGR gas in the mixed gas, is set so that when the injected fuel is burned, the temperature of the fuel and the surrounding gas is lower than the temperature at which soot is formed. The required minimum EGR gas amount is shown. This EGR gas amount is approximately 55% or more in terms of the EGR rate. In the embodiment shown in FIG.
0% or more. That is, the total intake gas amount sucked into the combustion chamber 5 is indicated by a solid line X in FIG. 6, and the ratio of the air amount to the EGR gas amount in the total intake gas amount X is shown in FIG.
When the ratio is as shown in the following, the temperature of the fuel and the surrounding gas is lower than the temperature at which soot is generated, and thus no soot is generated. Further, the NO _x generation amount at this time is around 10 p.pm or less.
The amount of O _x generated is extremely small.

When the fuel injection amount increases, the amount of heat generated when the fuel burns increases. Therefore, in order to maintain the temperature of the fuel and the surrounding gas at a temperature lower than the temperature at which the soot is generated, the heat generated by the EGR gas is required. Must be increased. Therefore, as shown in FIG. 6, the EGR gas amount must be increased as the injected fuel amount increases.
That is, the EGR gas amount needs to increase as the required load increases.

When the supercharging is not performed, the upper limit of the total intake gas amount X sucked into the combustion chamber 5 is Y. Therefore, in FIG. 6, the required load is larger in the region where the required load is larger than Lo. As the ratio increases, the air-fuel ratio cannot be maintained at the stoichiometric air-fuel ratio unless the EGR gas ratio is reduced. In other words, when the supercharging is not performed and the required air-fuel ratio is maintained at the stoichiometric air-fuel ratio in an area where the required load is larger than Lo, the EGR rate decreases as the required load increases, and In the region where the required load is larger than Lo, the temperature of the fuel and the surrounding gas cannot be maintained at a temperature lower than the temperature at which soot is generated.

However, as shown in FIG. 1, when the EGR gas is recirculated through the EGR passage 29 to the inlet side of the supercharger, that is, into the air suction pipe 17 of the exhaust turbocharger 15, the required load is larger than Lo. In EGR rate 5
It can be maintained at 5% or more, for example 70%, so that the temperature of the fuel and its surrounding gas can be kept below the temperature at which soot is produced. That is, if the EGR gas is recirculated so that the EGR rate in the air suction pipe 17 becomes, for example, 70%, the EGR rate of the suction gas boosted by the compressor 16 of the exhaust turbocharger 15 also becomes 70%. The temperature of the fuel and the surrounding gas can be maintained at a temperature lower than the temperature at which soot is generated, to the extent that the pressure can be increased by the compressor 16. Therefore, the operating range of the engine that can cause low-temperature combustion can be expanded.

In this case, when the EGR rate is set to 55% or more in a region where the required load is larger than Lo, the EGR control valve 31 is fully opened and the throttle valve 20
Is slightly closed. As described above, FIG. 6 shows the case where the fuel is burned under the stoichiometric air-fuel ratio. However, even if the air amount is smaller than the air amount shown in FIG. 10p.pm the generation amount of the NO _x while preventing the occurrence of longitudinal or can below, also be more than the amount of air shown the amount of air in FIG. 6, that is, the average of the air-fuel ratio Even when leaning from 17 to 18, the amount of generated NO _x can be reduced to around 10 ppm or less while preventing the generation of soot.

That is, when the air-fuel ratio is made rich, the fuel becomes excessive, but since the combustion temperature is suppressed to a low temperature, the excess fuel does not grow to soot, thus producing soot. There is no. Further, at this time NO _x even only an extremely small amount of generated. On the other hand, when the average air-fuel ratio is lean, or even when the air-fuel ratio is the stoichiometric air-fuel ratio, a small amount of soot is generated if the combustion temperature increases, but in the present invention, the soot is suppressed to a low temperature, so that the soot is reduced. Not generated at all. Furthermore, NO _x
Only very small amounts are generated.

As described above, when low-temperature combustion is performed, soot is generated regardless of the air-fuel ratio, that is, whether the air-fuel ratio is rich, the stoichiometric air-fuel ratio, or the average air-fuel ratio is lean. Sarezu, the amount of the NO _x becomes extremely small. Therefore, considering the improvement of the fuel consumption rate, it can be said that it is preferable to make the average air-fuel ratio lean at this time.

By the way, the temperature of the fuel during combustion in the combustion chamber and the gas around it can be suppressed to a temperature below the temperature at which the growth of hydrocarbons stops halfway, only when the engine is operating at a low load and the calorific value due to combustion is relatively small. Can be Therefore, in the embodiment according to the present invention, the first combustion, that is, the low-temperature combustion is performed by suppressing the temperature of the fuel during combustion and the surrounding gas temperature to a temperature at which the growth of hydrocarbons stops halfway during the low load operation in the engine. During the high load operation of the engine, the second combustion, that is, the combustion that is usually performed conventionally, is performed. Here, the first combustion, that is, the low-temperature combustion, has a larger amount of the inert gas in the combustion chamber than the amount of the inert gas at which the soot generation amount is at a peak, as is clear from the description so far. The second combustion, that is, the combustion that has been performed normally in the past, is a combustion in which the amount of inert gas in the combustion chamber is smaller than the amount of inert gas at which the amount of soot is peaked. Say that.

FIG. 7 shows a first operation region I in which the first combustion, that is, low-temperature combustion is performed, and a second operation region II in which the second combustion, that is, combustion by the conventional combustion method, is performed. I have. In FIG. 7, the vertical axis L indicates the amount of depression of the accelerator pedal 50, that is, the required load, and the horizontal axis N indicates the engine speed. In FIG. 7, X (N) is the first
Shows the first boundary between the operating region I and the second operating region II, and Y (N) represents the first operating region I and the second operating region.
2 shows a second boundary with II. The determination of the change of the operating region from the first operating region I to the second operating region II is made based on the first boundary X (N), and the determination of the change from the second operating region II to the first operating region II is performed.
The determination of the change of the operation region to the operation region I of the second boundary Y
(N).

That is, when the operating state of the engine is in the first operating region I
When the required load L exceeds a first boundary X (N), which is a function of the engine speed N, during low-temperature combustion, it is determined that the operation region has shifted to the second operation region II, Combustion is performed by a conventional combustion method. Next, when the required load L becomes lower than a second boundary Y (N) which is a function of the engine speed N, it is determined that the operation region has shifted to the first operation region I, and low-temperature combustion is performed again.

The two boundaries of the first boundary X (N) and the second boundary Y (N) on the lower load side than the first boundary X (N) are provided as follows. For three reasons. The first reason is that the combustion temperature is relatively high on the high load side of the second operation region II, and even if the required load L becomes lower than the first boundary X (N), low-temperature combustion cannot be performed immediately. Because. That is, the low-temperature combustion does not immediately start unless the required load L becomes considerably low, that is, when the required load L becomes lower than the second boundary Y (N). The second reason is to provide a hysteresis for a change in the operation region between the first operation region I and the second operation region II.

By the way, when the operating state of the engine is in the first operating region I and the low-temperature combustion is being performed, soot is hardly generated, but the unburned hydrocarbon is replaced with the precursor of soot or the state before the soot. It is discharged from the combustion chamber 5 in the form. At this time, the unburned hydrocarbon discharged from the combustion chamber 5 is oxidized well by the catalyst 25 having an oxidizing function. Catalyst 2
As 5, an oxidation catalyst, a three-way catalyst, or a NO _x absorbent can be used. _The NO _x absorbent absorbs NO _x when the mean air-fuel ratio in the combustion chamber 5 of the lean, the combustion chamber 5
The average air-fuel ratio in the internal has a function of releasing NO _x becomes rich.

This NO _x absorbent uses, for example, alumina as a carrier and, for example, potassium K, sodium N
a, at least one selected from alkali metals such as lithium Li, cesium Cs, alkaline earths such as barium Ba and calcium Ca, rare earths such as lanthanum La and yttrium Y, and a noble metal such as platinum Pt. Is carried.

In addition to the oxidation catalyst, the three-way catalyst and the NO
_{The x} absorbent also has an oxidizing function, so that a three-way catalyst and a NO _x absorbent can be used as the catalyst 25 as described above. FIG. 8 shows the output of the air-fuel ratio sensor 27. As shown in FIG. 8, the output current I of the air-fuel ratio sensor 27 changes according to the air-fuel ratio A / F. Therefore, the air-fuel ratio can be known from the output current I of the air-fuel ratio sensor 27.

Next, the operation control in the first operation region I and the second operation region II will be schematically described with reference to FIG. FIG. 9 shows the throttle valve 2 with respect to the required load L.
0 indicates the opening degree, the opening degree of the EGR control valve 31, the EGR rate, the air-fuel ratio, the injection timing, and the injection amount. As shown in FIG. 9, in the first operating region I where the required load L is low, the opening of the throttle valve 20 is gradually increased from almost fully closed to about 2/3 opening as the required load L increases. E
The degree of opening of the GR control valve 31 is gradually increased from almost fully closed to fully open as the required load L increases. Also,
In the example shown in FIG. 9, in the first operation region I, the EGR rate is set to approximately 70%, and the air-fuel ratio is set to a slightly lean air-fuel ratio.

In other words, in the first operating region I, the EGR
The opening of the throttle valve 20 and the opening of the EGR control valve 31 are controlled such that the rate becomes approximately 70% and the air-fuel ratio becomes a slightly lean air-fuel ratio. At this time, the air-fuel ratio is controlled to the target lean air-fuel ratio by correcting the opening of the EGR control valve 31 based on the output signal of the air-fuel ratio sensor 27. In the first operation region I, fuel injection is performed before the compression top dead center TDC. In this case, the injection start timing θ is delayed as the required load L increases.

During the idling operation, the throttle valve 20 is closed until the valve is almost fully closed. At this time, the EGR control valve 31 is also closed almost completely. Throttle valve 2
If the valve is closed close to 0, the pressure in the combustion chamber 5 at the start of compression decreases, and the compression pressure decreases. When the compression pressure decreases, the compression work by the piston 4 decreases, so that the vibration of the engine body 1 decreases. That is, at the time of idling operation, the throttle valve 20 is closed to almost fully closed in order to suppress the vibration of the engine body 1.

On the other hand, the operating range of the engine is the first operating range I.
From the second operating region II to the second operating region II, the opening of the throttle valve 20 is increased stepwise from about 2/3 opening toward the full opening direction. At this time, in the example shown in FIG. 9, the EGR rate is reduced stepwise from approximately 70% to 40% or less so that the EGR rate jumps over the EGR rate range (FIG. 5) where a large amount of smoke is generated, and the air-fuel ratio is reduced. It is enlarged in steps.

In the second operation region II, the second combustion, that is, the conventional combustion is performed. In this combustion method generates little soot and NO _x, but the heat efficiency is higher than the low temperature combustion, thus as the operating region of the engine is shown in Figure 9 from the first operation area I changes to the second operating region II Thus, the injection amount is reduced stepwise. Second operating area
In II, conventional combustion is performed. This second
In the operating region II, the throttle valve 20 is held in a fully open state except for a part, and the opening degree of the EGR control valve 31 is equal to the required load L.
Are gradually reduced as they become higher. In addition, this operating area
In II, the EGR rate decreases as the required load L increases,
The air-fuel ratio decreases as the required load L increases. However, the air-fuel ratio is a lean air-fuel ratio even when the required load L increases. In the second operation region II, pilot injection Qp is performed prior to main injection Qm.

FIG. 10 shows the air-fuel ratio A / F in the first operating region I. In FIG. 10, A / F = 15.
The curves indicated by 5, A / F = 16, A / F = 17, and A / F = 18 have air-fuel ratios of 15.5, 16, 17, and 18, respectively.
And the air-fuel ratio between the curves is determined by proportional distribution. As shown in FIG. 10, the air-fuel ratio is lean in the first operating region I, and in the first operating region I, the air-fuel ratio A / F becomes leaner as the required load L decreases.

That is, the lower the required load L, the smaller the amount of heat generated by combustion. Therefore, low-temperature combustion can be performed even if the EGR rate is reduced as the required load L decreases.
When the EGR rate is reduced, the air-fuel ratio increases. Therefore, as shown in FIG. 10, as the required load L decreases, the air-fuel ratio A / F increases. As the air-fuel ratio A / F increases, the fuel consumption rate increases. Accordingly, in order to make the air-fuel ratio as lean as possible, in the embodiment according to the present invention, the air-fuel ratio A / F increases as the required load L decreases.

The target opening degree ST of the throttle valve 20 required for setting the air-fuel ratio to the target air-fuel ratio shown in FIG.
As shown in FIG. 1A, the EGR is stored in the ROM 42 in advance in the form of a map as a function of the required load L and the engine speed N, and is necessary for setting the air-fuel ratio to the target air-fuel ratio shown in FIG. The target opening SE of the control valve 31 is as shown in FIG.
As shown in (1), a map is previously stored in the ROM 42 as a function of the required load L and the engine speed N.

FIG. 12 shows the target air-fuel ratio when the second combustion, that is, the combustion by the conventional combustion method is performed.
In FIG. 12, A / F = 24, A / F = 35, A
Each curve indicated by / F = 45 and A / F = 60 indicates target air-fuel ratios 24, 35, 45, and 60, respectively. As shown in FIG. 13A, the target opening ST of the throttle valve 20 required for setting the air-fuel ratio to the target air-fuel ratio is equal to the required load L as shown in FIG.
And RO in the form of a map as a function of the engine speed N
The target opening SE of the EGR control valve 31 which is stored in the M42 and is necessary for setting the air-fuel ratio to the target air-fuel ratio is shown in FIG.
As shown in FIG. 3 (B), it is stored in the ROM 42 in advance in the form of a map as a function of the required load L and the engine speed N.

As shown in FIG. 9, when the first combustion is performed in the first operation region I, the fuel injection Q is performed only once at the end of the compression stroke, and the second injection is performed in the second operation region II. As described above, the pilot injection Qp is performed prior to the main injection Qm when the combustion is performed. As shown in FIG. 9, the injection amount of the fuel injection Q, the injection amount of the main injection Qm, and the injection amount of the pilot injection Qp increase as the required load L increases.

Actually, the injection amount of the fuel injection Q is equal to the required load L
And the engine speed N, the injection amount Q
As shown in FIG. 4A, a map is stored in advance in the ROM 42 as a function of the required load L and the engine speed N. Similarly, the injection amount of the main injection Qm is also a function of the required load L and the engine speed.
As shown in (B), the required load L and the engine speed N
Is stored in the ROM 42 in advance in the form of a map as a function of. Further, the injection quantity of the pilot injection Qp is also a function of the required load L and the engine speed N. This injection quantity Qp is also a function of the required load L and the engine speed N in the map as shown in FIG. The information is stored in the ROM 42 in advance in the form.

FIG. 15A shows the injection start timing of the fuel injection Q and the injection start timing of the main injection Qm. That is, in FIG. 15 (A), the load side lower than the first boundary X (N) indicates the injection start timing of the fuel injection Q.
The load higher than the boundary X (N) indicates the injection start timing of the main injection Qm. In FIG. 15A, each numerical value indicates the crank angle of BTDC before top dead center, and each solid line indicates the equal injection start timing.

Since the first combustion, that is, the low-temperature combustion is performed in a state where the EGR rate is high, the combustion is slow. Therefore, if the injection start timing is delayed at this time, such as the injection start timing of the main injection Qm, good combustion is performed. Failure to do so will result in misfires. Therefore, as can be seen from FIG. 15 (A), the injection start timing under the first combustion is earlier than the injection start timing of the main injection Qm under the second combustion.

The same applies to the case where the fuel injection is performed only once in the second combustion. In this case, the injection start timing in the first combustion is the same as that in the second combustion. It is earlier than time. Therefore the first
When the combustion is switched to the second combustion, the injection start timing is delayed, and when the combustion is switched from the second combustion to the first combustion, the injection start timing is advanced.

As can be seen from FIG. 15 (A), the injection start timing during the first combustion is a function of the required load L and the engine speed N.
As shown in FIG. 5 (B), a map is stored in advance in the ROM 42 as a function of the required load L and the engine speed N. As can be seen from FIG. 15A, the injection start timing of the main injection Qm is also a function of the required load L and the engine speed N. The injection start timing .theta.m is also the required load as shown in FIG. It is stored in the ROM 42 in advance in the form of a map as a function of L and the engine speed N.

FIG. 16A shows a change in the injection start timing θm of the main injection Qm and a change in the injection start timing θp of the pilot injection Qp when the required load L changes under a constant engine speed N. FIG. 16B shows the main injection Qm when the engine speed N changes with the required load L constant.
And the change of the injection start timing θp of the pilot injection Qp. As can be seen from FIGS. 16A and 16B, pilot injection Qp
The injection start timing θp is a function of the required load L and the engine speed N, and the injection start timing θp is a function of the required load L and the engine speed N in the form of a map as shown in FIG. It is stored in the ROM 42 in advance.

Next, the control of the injection amount and the injection start timing when the first combustion is switched to the second combustion will be described with reference to FIGS. 17 (A) and 17 (B). As described above, when the first combustion is switched to the second combustion, the EGR is performed by opening the throttle valve 20.
The rate is reduced, for example, from 70 percent to less than 40 percent. However, at this time, the EGR rate does not decrease instantaneously due to a delay in increasing the intake air amount or the like, and a certain time is required until the EGR rate decreases from 70% to 40% or less. That is, when the first combustion is switched to the second combustion, the EGR rate suitable for the first combustion gradually changes from the EGR rate suitable for the second combustion. Therefore, when the first combustion is switched to the second combustion, the injection amount and the injection start timing are changed from the injection amount and the injection start timing suitable for the first combustion to the injection amount and the injection start timing suitable for the second combustion. It would be preferable to change it gradually.

In the embodiment according to the present invention, FIG.
17A and 17B, when the first combustion is switched to the second combustion, the main injection amount Qmi is gradually increased, and the pilot injection amount Qpi is gradually decreased. The injection start timing θmi of the injection and the injection start timing θpi of the pilot injection are gradually changed.
Note that FIGS. 17A and 17B show t ₁ to t _4.
The change of the injection period when the time elapses toward is shown. FIG. 17A shows a case where the engine speed N is relatively low, and FIG. 17B shows a case where the engine speed N is relatively high. FIG. 17 (A) and FIG.
In (B), the horizontal axis indicates the crank angle. Therefore, even if the injection amount is the same, the injection period becomes longer as the engine speed N increases.

In FIGS. 17 (A) and 17 (B), t ₁ indicates the injection period when the first combustion is being performed, and t ₂ and t ₃ indicate the first combustion to the second combustion. This shows a stage in the middle of the transition to combustion, and t ₄ shows a time when the second combustion is completely switched. FIG. 17 (A) and FIG.
As shown in FIG. 7B, when the first combustion is switched to the second combustion, the main injection amount Qmi is gradually increased toward the final main injection amount Qm, and the pilot injection amount Qpi is changed to the first injection amount. Is gradually decreased from the injection amount Q during the combustion toward the final pilot injection amount Qp. Note that at t ₂ and t ₃ , Qmi is called the main injection, and Qpi
Although it is not necessarily technically appropriate to refer to as pilot injection, in the specification of the present application, the injection performed earlier in the middle of the transition from the first combustion to the second combustion will be described in order to facilitate understanding of the invention. This is referred to as pilot injection, and the injection performed later is referred to as main injection.

On the other hand, when switching from the first combustion to the second combustion, the injection start timing θmi of the main injection is gradually advanced toward the final injection start timing θm. Also, the first
When the combustion is switched from the first combustion to the second combustion, the injection start timing θpi of the pilot injection is gradually changed from the injection start timing θ under the first combustion to the final injection start timing θp under the second combustion. Can be

When the engine speed N is low, FIG.
As shown in FIG. 17A, the injection start timing θ under the first combustion is relatively late.
As shown in (2), the injection start timing θpi of the pilot injection is gradually advanced. On the other hand, when the engine speed N is high, the injection start timing θ under the first combustion is relatively early as shown in FIG. 15A, and therefore, at this time, as shown in FIG. The injection start timing θpi of the pilot injection is gradually delayed.

When switching from the second combustion to the first combustion, the injection period is changed from t ₄ to t ₁ in FIGS. 17 (A) and 17 (B). Therefore, when switching from the second combustion to the first combustion, the main injection amount Qmi gradually decreases, the pilot injection amount Qpi increases gradually, and the main injection start timing θmi and the pilot injection start timing θpi will be gradually changed.

Next, the operation control will be described with reference to FIG. Referring to FIG. 18, first, Step 1
At 00, it is determined whether or not a flag I indicating that the operating state of the engine is in the first operating region I is set. When the flag I is set, that is, when the operating state of the engine is in the first operating region I, step 10
The program proceeds to 1 to determine whether the required load L has become larger than the first boundary X1 (N). When L ≦ X1 (N), the routine proceeds to step 104, where low-temperature combustion is performed.

That is, in step 104, the target opening ST of the throttle valve 20 is calculated from the map shown in FIG. 11A, and the opening of the throttle valve 20 is set to the target opening ST. Next, at step 105, the target opening degree SE of the EGR control valve 31 is calculated from the map shown in FIG.
The opening of the EGR control valve 31 is set as the target opening SE.
Next, at step 106, the injection control I shown in FIG.
Is performed.

On the other hand, in step 101, L> X
When it is determined that (N) has been reached, the routine proceeds to step 102, where the flag I is reset.
Then, the completion flag indicating that the injection control at the time of switching between the first combustion and the second combustion is completed is reset. Next, the routine proceeds to step 110, where the second combustion is performed.

That is, in step 110, the target opening ST of the throttle valve 20 is calculated from the map shown in FIG. 13A, and the opening of the throttle valve 20 is set to the target opening ST. Next, at step 111, the target opening SE of the EGR control valve 31 is calculated from the map shown in FIG.
The opening of the EGR control valve 31 is set as the target opening SE.
Next, at step 112, the injection control II shown in FIG.
Is performed.

When the flag I is reset, in the next processing cycle, the process proceeds from step 100 to step 107, where it is determined whether or not the required load L has become lower than the second boundary Y (N). Step 110 when L ≧ Y (N)
And the second combustion is performed. On the other hand, step 107
When it is determined that L <Y (N), the routine proceeds to step 108, where the flag I is set. Then, the routine proceeds to step 109, where the completion flag is reset. Next, the routine proceeds to step 104, where low-temperature combustion is performed.

Next, the injection control I will be described with reference to FIGS. 19 and 20. Referring to FIGS. 19 and 20, first, at step 200, it is determined whether or not the completion flag is set. Since the completion flag is reset when the second combustion is switched to the first combustion, the process proceeds to step 201 when the second combustion is switched to the first combustion.

In step 201, the main injection amount Qm immediately before switching from the second combustion to the first combustion is set to a constant integer value I
The result is divided by NT, and the result of the division is set to ΔQm. Then delta Q m is subtracted from the main injection amount Qmi step 202. Next, at step 203, it is determined whether or not the main injection amount Qmi has become smaller than the minimum value MIN. Qmi <
When MIN is reached, the routine proceeds to step 204, where the main injection amount Qmi is made zero. Therefore, it is understood that the main injection amount Qmi is gradually reduced when the mode is switched from the second combustion to the first combustion.

Next, at step 205, a constant value α is added to Δθm. Next, at step 206, .DELTA..theta.m is subtracted from the injection start timing .theta.m of the main injection immediately before switching from the second combustion to the first combustion, and the subtraction result is the injection start timing .theta.mi of the main injection. Next, at step 207, it is determined whether or not the injection start timing θmi of the main injection has become smaller than a predetermined injection start timing R, that is, whether or not it has become late. When θmi <R, the routine proceeds to step 208, where θmi is set to R. Therefore, it can be seen that when the second combustion is switched to the first combustion, the injection start timing θmi of the main injection is gradually delayed.

Next, at step 209, the difference (Q-Qp) between the injection amount Q under the first combustion and the pilot injection amount Qp immediately before switching from the second combustion to the first combustion.
Is divided by a constant integer INT, and the result of the division is ΔQp
It is said. Next, at step 210, the integrated value S is added to ΔQp
Is added. Next, at step 211, the integrated value S is added to the pilot injection amount Qp, and the addition result is used as the pilot injection amount Qpi. Next, at step 212, it is determined whether or not the pilot injection amount Qpi has become larger than the injection amount Q under the first combustion. When Qpi> Q, the routine proceeds to step 213, where the pilot injection amount Qpi is set to Q. Therefore, it can be seen that the pilot injection amount Qpi is gradually increased when switching from the second combustion to the first combustion.

Next, at step 214, the injection start timing θp of the pilot injection immediately before switching from the second combustion to the first combustion, and the injection start timing θ under the first combustion
(Θp−θ) is divided by a constant integer value INT, and the result of the division is Δθp. Then step 215
Then, Δθp is added to the integrated value SS. Next, at step 216, the integrated value SS is subtracted from the injection start timing θp of the pilot injection, and the subtraction result is set as the injection start timing θpi of the pilot injection. Next, at step 217, it is determined whether or not the absolute value | θpi−θ | of the difference between the injection start timing θpi of the pilot injection and the injection start timing θ under the first combustion has become smaller than the fixed value β. −θ | <β
When, the routine proceeds to step 218, where the injection start timing θpi of the pilot injection is set to θ. Therefore, it is understood that when the second combustion is switched to the first combustion, the injection start timing θpi of the pilot injection is gradually changed.

Next, at step 219, step 204
At step 208, Qmi is made zero, and at step 208, θmi
Is set to R, and Qpi is set to Q in step 213.
In step 218, it is determined whether or not θpi has been set to θ, that is, whether or not switching to the first combustion has been completed.
When the switching to the first combustion is completed, step 22 is executed.
Go to 0, complete flag is set, then step 2
At 21, Qmi, Δθm, S and SS are made zero.

When the completion flag is set, step 20
From 0, the routine proceeds to step 222, where the injection amount Q is calculated from the map shown in FIG. 14A, and then to step 223, where the injection start timing θ is calculated from the map shown in FIG. 15B. Therefore, at this time, the first combustion is performed. Next, the injection control II will be described with reference to FIGS.

Referring to FIGS. 21 and 22, first, at step 300, it is determined whether or not the completion flag is set. Since the completion flag is reset when the first combustion is switched to the second combustion, the process proceeds to step 301 when the first combustion is switched to the second combustion. In step 301, the main injection amount Qm immediately after switching from the first combustion to the second combustion is divided by a constant integer value INT, and the division result is set as ΔQm. Then delta Q m is added to the main injection amount Qmi step 302. Next, at step 303, it is determined whether or not the main injection amount Qmi has become larger than the injection amount Qm under the second combustion. Step 304 when Qmi> Qm
The main injection amount Qmi is set to Qm. Therefore, it can be seen that when the first combustion is switched to the second combustion, the main injection amount Qmi is gradually increased.

Next, at step 305, a constant value α is added to Δθm. Next, at step 306, a predetermined period T is subtracted from the injection start timing θm of the main injection immediately after switching from the first combustion to the second combustion, and Δθm is added. The injection start timing is θmi. Next, at step 307, it is determined whether the injection start timing θmi of the main injection has become larger than the injection start timing θm of the main injection under the second combustion, that is, whether it has been advanced. When θmi> θm, step 30
Proceeding to 8, θmi is set to θm. Therefore, when the first combustion is switched to the second combustion, the injection start timing θmi of the main injection
It can be seen that is gradually accelerated.

Next, at step 309, the difference (Q-Qp) between the injection amount Q under the first combustion and the pilot injection amount Qp immediately after switching from the first combustion to the second combustion.
Is divided by a constant integer INT, and the result of the division is ΔQp
It is said. Next, at step 310, the integrated value S is added to ΔQp
Is added. Next, at step 311, the integrated value S is subtracted from the pilot injection amount Qp, and the subtraction result is used as the pilot injection amount Qpi. Next, at step 312, it is determined whether or not the pilot injection amount Qpi has become smaller than the pilot injection amount Qp under the second combustion. Qpi
When <Qp, the routine proceeds to step 313, where the pilot injection amount Qpi is set to Qp. Therefore, it can be seen that when the first combustion is switched to the second combustion, the pilot injection amount Qpi is gradually reduced.

Next, at step 314, the injection start timing θp of the pilot injection immediately after switching from the first combustion to the second combustion, and the injection start timing θ under the first combustion
(Θp−θ) is divided by a constant integer value INT, and the result of the division is Δθp. Then step 315
Then, Δθp is added to the integrated value SS. Next, at step 316, the integrated value SS is added to the injection start timing θ under the first combustion, and the addition result is used as the pilot start injection start timing θpi. Next, at step 317, the absolute value | θpi of the difference between the injection start timing θpi of the pilot injection and the injection start timing θp of the pilot injection under the second combustion.
It is determined whether or not -.theta.p | has become smaller than the fixed value .beta .. When | .theta.pi-.theta.p | <.beta.
Proceeding to 8, the injection start timing θpi of the pilot injection is set to θp. Therefore, it is understood that when the first combustion is switched to the second combustion, the injection start timing θpi of the pilot injection is gradually changed.

Next, at step 319, step 304 is executed.
In step 308, Qmi is set to Qm.
mi is set to θm, Qpi is set to Qp in step 313, and it is determined in step 318 whether or not θpi is set to θp, that is, whether or not the switching to the second combustion is completed. When the switching to the second combustion is completed, the routine proceeds to step 320, where a completion flag is set. Then, in step 321, Qmi, Δθm, S, and SS are made zero.

When the completion flag is set, step 30
From 0, the routine proceeds to step 322, where the main injection amount Qm is calculated from the map shown in FIG.
The injection start timing θm of the main injection is calculated from the map shown in FIG. Next, the routine proceeds to step 324, where the pilot injection amount Qp is obtained from the map shown in FIG.
Is calculated, and then the routine proceeds to step 325, where FIG.
The injection start timing θp of the pilot injection is calculated from the map shown in FIG. Therefore, at this time, the second combustion is performed.

[0096]

The first combustion and the second combustion can be properly switched.

[Brief description of the drawings]

FIG. 1 is an overall view of a compression ignition type internal combustion engine.

FIG. 2 is a diagram showing amounts of smoke and NO _x generated, and the like.

FIG. 3 is a diagram showing a combustion pressure.

FIG. 4 is a diagram showing fuel molecules.

FIG. 5 is a diagram showing a relationship between a generation amount of smoke and an EGR rate.

FIG. 6 is a diagram showing a relationship between an injected fuel amount and a mixed gas amount.

FIG. 7 is a diagram showing a first operation region I and a second operation region II.

FIG. 8 is a diagram showing an output of an air-fuel ratio sensor.

FIG. 9 is a diagram showing an opening degree of a throttle valve and the like.

FIG. 10 is a diagram showing an air-fuel ratio in a first operation region I.

FIG. 11 is a diagram showing a map of a target opening degree of a throttle valve and the like.

FIG. 12 is a diagram showing an air-fuel ratio in a second combustion.

FIG. 13 is a view showing a map of a target opening degree of a throttle valve and the like.

FIG. 14 is a diagram showing a map of an injection amount.

FIG. 15 is a diagram showing an injection start timing.

FIG. 16 is a diagram showing pilot injection timing.

FIG. 17 is a diagram showing an injection period.

FIG. 18 is a flowchart for controlling operation of the engine.

FIG. 19 is a flowchart for executing an injection control I.

FIG. 20 is a flowchart for executing injection control I;

FIG. 21 is a flowchart for executing injection control II.

FIG. 22 is a flowchart for executing injection control II.

[Explanation of symbols]

 6: fuel injection valve 15: exhaust turbocharger 20: throttle valve 29: EGR passage

────────────────────────────────────────────────── ─── front page continued (51) Int.Cl. ⁷ identifications FI F02D 43/00 301 F02D 43/00 301H 301J 301N F02M 25/07 550 F02M 25/07 550R 570 570J (72) inventor Masato Goto Aichi 1 Toyota Town, Toyota City, Toyota Prefecture, Japan (72) Inventor Hiroki Murata 1 Toyota Town, Toyota City, Aichi Prefecture, Toyota Motor Corporation (56) References JP-A-11-22535 (JP, A) JP-A-11-154949 (JP, A) JP-A-11-107820 (JP, A) JP-A-8-303309 (JP, A) JP-A-9-287528 (JP, A) (58) (Int.Cl. ⁷ , DB name) F02D 41/00-41/40 F01N 3/08-3/28 F02D 43/00 301 F02M 25/07 550 F02M 25/07 570

Claims (13)

(57) [Claims] When the amount of inert gas in the combustion chamber increases, the amount of soot generated gradually increases and reaches a peak. When the amount of inert gas in the combustion chamber further increases, the amount of soot generated during combustion in the combustion chamber increases. In the internal combustion engine where the temperature of the fuel and the surrounding gas is lower than the soot generation temperature and soot is hardly generated, the fuel injection valve is arranged in the combustion chamber, and the amount of inert gas at which the amount of soot generation peaks The first combustion in which the amount of inert gas in the combustion chamber is large and soot is hardly generated, and the second combustion in which the amount of inert gas in the combustion chamber is smaller than the amount of inert gas at which the amount of generated soot reaches a peak. Switching means for selectively switching between the first combustion and the second combustion, the fuel injection start timing is delayed, and the fuel injection start timing is switched when the second combustion is switched to the first combustion. To speed up Internal combustion engine. 2. At least at the time of switching between the first combustion and the second combustion, the second combustion is performed by performing the pilot injection prior to the main injection, and the first combustion is switched to the second combustion. When the first combustion is performed, the injection start timing of the main injection is delayed as compared with the injection start timing when the first combustion is performed. When the second combustion is switched to the first combustion, the injection start timing of the main injection is performed. 2. The internal combustion engine according to claim 1, wherein the injection start timing is advanced as compared with the internal combustion engine. 3. The internal combustion engine according to claim 2, wherein when the first combustion is switched to the second combustion, the injection amount of the main injection is gradually increased and the injection amount of the pilot injection is gradually decreased. 4. When the first combustion is switched to the second combustion, the injection start timing of the pilot injection is predetermined according to the operating state of the engine during the second combustion from the injection start timing during the first combustion. 4. The internal combustion engine according to claim 3, wherein the internal combustion engine is gradually changed toward the pilot injection start timing. 5. The internal combustion engine according to claim 3, wherein when the first combustion is switched to the second combustion, the injection start timing of the main injection is gradually advanced. 6. The internal combustion engine according to claim 2, wherein when the second combustion is switched to the first combustion, the injection amount of the main injection is gradually reduced and the injection amount of the pilot injection is gradually increased. 7. The internal combustion engine according to claim 6, wherein when the second combustion is switched to the first combustion, the injection start timing of the pilot injection is gradually changed toward the injection start timing in the first combustion. organ. 8. The internal combustion engine according to claim 6, wherein when the second combustion is switched to the first combustion, the injection start timing of the main injection is gradually delayed. 9. The internal combustion engine according to claim 1, further comprising a recirculation device for recirculating exhaust gas discharged from the combustion chamber into the engine intake passage, wherein the inert gas comprises recirculated exhaust gas. 10. The internal combustion engine according to claim 9, wherein the exhaust gas recirculation rate in the first combustion state is approximately 55% or more. 11. The internal combustion engine according to claim 1, wherein a catalyst having an oxidation function is disposed in the engine exhaust passage. 12. The catalyst according to claim 1, wherein said catalyst is an oxidation catalyst, a three-way catalyst or NO.
An internal combustion engine according to claim 11, comprising at least one of the _x absorbents. 13. An engine operating region is divided into a first operating region on a low load side and a second operating region on a high load side, and a first combustion is performed in the first operating region, and a second operation is performed. Second in the area
The internal combustion engine according to claim 1, wherein combustion of the internal combustion engine is performed.