JP3405221B2 - Internal combustion engine - Google Patents

Description

Detailed Description of the Invention

[0001]

TECHNICAL FIELD The present invention relates to an internal combustion engine.

[0002]

2. Description of the Related Art Conventionally, in an internal combustion engine, for example, a diesel engine, an engine exhaust passage and an engine intake passage are connected by an exhaust gas recirculation (hereinafter referred to as EGR) passage in order to suppress the generation of NO _X. Exhaust gas, that is, EGR gas, is recirculated into the engine intake passage via the EGR 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 decreases, the NO _X generation amount decreases, and therefore, the higher the EGR rate, the lower the NO _X generation amount.

As described above, it has been known that the amount of NO _x generated can be reduced by increasing the EGR rate. However, when the EGR rate is increased, when the EGR rate exceeds a certain limit, the amount of soot generated, that is, the smoke starts to increase rapidly. In this regard, it has been conventionally thought that if the EGR rate is further increased, the smoke will increase infinitely, and therefore the smoke will start to increase rapidly.
The GR rate is considered to be the maximum allowable limit for the EGR rate.

Therefore, conventionally, the EGR rate is set within a range that does not exceed the maximum allowable limit. The maximum allowable limit of this EGR rate is approximately 30 to 50 percent, though it varies considerably depending on the engine type and fuel.
Therefore, in the conventional diesel engine, the maximum EGR rate is 3
It is suppressed from 0% to 50%.

As described above, in the past, it was considered that the maximum allowable limit exists for the EGR rate.
The R rate is NO within the range that does not exceed this maximum allowable limit.
_It was set to minimize the amount of _X and smoke generated. However, in this way EGR
Even if the rate is set so that the amount of NO _x and smoke produced is as small as possible, there is a limit to the reduction in the amount of NO _x and smoke produced, and in reality, a considerable amount of N 2 still remains.
The O _X and the smoke is generated at present.

However, if the EGR rate is made larger than the maximum permissible limit in the process of studying the combustion of a diesel engine, the smoke increases sharply as described above, but there is a peak in the amount of smoke produced, and the peak is exceeded. EGR
When the rate is further increased, the smoke starts to decrease sharply this time. When the EGR rate is 70% or more during idling operation, and when the EGR gas is strongly cooled, the EGR rate is almost 55% or more. It was found that it was almost zero, that is, soot was hardly generated. At this time, N
Generation amount of O _X is also found that a very small amount.
After that, the reason why soot was not generated was examined based on this finding, and as a result, soot and NO
_This led to the construction of a new combustion system capable of simultaneously reducing _X. This new combustion system will be explained in detail later, but in short, it is basically based on stopping the growth of hydrocarbons in the middle of the process until the hydrocarbons grow into soot.

[0007] That is, as a result of repeated experimental research, it was found that when the temperature of the fuel and the gas around it during combustion in the combustion chamber were below a certain temperature, the growth of hydrocarbons stopped in the middle of the process before reaching soot. However, if the temperature of the fuel and the gas around it rises above a certain temperature, the hydrocarbons will suddenly grow to soot. In this case, the temperature of the fuel and its surrounding gas is greatly affected by the endothermic action of the gas around the fuel when the fuel burns, and the endothermic amount of the gas around the fuel is adjusted according to the amount of heat generated during fuel combustion. Thus, the temperature of the fuel and the gas around it can be controlled.

Therefore, if the temperature of the fuel and the gas around it during combustion in the combustion chamber is suppressed below the temperature at which the growth of hydrocarbons stops halfway, soot will not be generated and the fuel and the surroundings during combustion in the combustion chamber will be eliminated. It is possible to control the gas temperature in the range below the temperature at which the growth of hydrocarbons stops halfway by adjusting the heat absorption amount of the gas around the fuel. On the other hand, hydrocarbons whose growth has stopped before reaching soot can be easily purified by post-treatment using an oxidation catalyst or the like. This is the basic idea of the new combustion system. The applicant has already applied for an internal combustion engine that employs this new combustion system (Japanese Patent Application No. 9-305850).

[0009]

By the way, in this new combustion system, it is necessary to increase the EGR rate to about 55% or more. The EGR rate can be set to about 55% or more when the intake air amount is relatively small. Is. That is, when the intake air amount exceeds a certain amount, this new combustion cannot be performed. Therefore, when the intake air amount exceeds a certain amount, the conventional combustion is switched to.

By the way, the new combustion has a slightly lower thermal efficiency than the conventional combustion, and therefore, the fuel injection amount required to generate the same engine output torque is higher in the new combustion than in the conventional combustion. It is a little more than the combustion. Therefore, when the combustion is switched from the new combustion to the conventional combustion, the fuel injection amount is reduced.

On the other hand, in the compression ignition type internal combustion engine, conventionally, the fuel injection amount is calculated from the depression amount of the accelerator pedal and the engine speed. Specifically, the fuel injection amount is stored in advance in the form of a map as a function of the accelerator pedal depression amount and the engine speed, and the fuel injection amount is calculated from this map. However, calculating the fuel injection amount using such a map causes a problem when the fuel injection amount needs to be reduced as in the case of switching from new combustion to conventional combustion as described above. Next, this will be described with reference to FIG.

In FIG. 19, L is the accelerator pedal depression amount, N is the engine speed, X is a new combustion operating region I and a conventional combustion operating region.
The boundaries with II are shown respectively. In this case, as shown in FIG. 19, the fuel injection amount Q1 for the operating region I is stored in the form of a map as a function of the accelerator pedal depression amount L and the engine speed N, and the fuel injection amount Q2 for the operating region II is stored. However, if it is stored in the form of a map as a function of the accelerator pedal depression amount L and the engine speed N, setting the fuel injection amount in the vicinity of the boundary X involves considerable difficulty.

That is, the most important thing is to prevent a torque step from occurring when the required load L exceeds the boundary X. For this purpose, the fuel injection amount Q1 and the fuel at all points on the boundary X are set. The injection amount Q2 and the injection amount Q2 must be set so that the same torque is generated at all points on the boundary X. However, in order to obtain the fuel injection amounts Q1 and Q2 at which the torque becomes the same at each point on the boundary X, it is necessary to determine the fuel injection amounts Q1 and Q2 at each point on the boundary X based on an experiment. Fuel injection amount Q1 on X,
There will be considerable difficulty in finding Q2.

[0014]

In order to solve the above problems, in the first invention, as the amount of inert gas in the combustion chamber is increased, the amount of soot generated gradually increases and reaches a peak.
When the amount of inert gas in the combustion chamber is further increased, the temperature of the fuel during combustion in the combustion chamber and the gas temperature around it become lower than the temperature at which soot is generated, and soot is hardly generated The amount of inert gas in the combustion chamber is larger than the peak amount of soot, and soot is rarely generated, and the amount of inert gas in the combustion chamber is more inert than the peak amount of soot generation. The first combustion and the second combustion are provided with a switching unit that selectively switches between the second combustion with a small amount of gas and a calculation unit that calculates the required torque of the engine.
The fuel injection amount is calculated on the basis of the calculated required torque and the engine speed regardless of which combustion is being performed. That is, the required torque is first obtained, and the fuel injection amount is calculated from the required torque and the engine speed.

In the second invention, in the first invention,
The calculating means calculates the required torque based on the depression amount of the accelerator pedal and the engine speed. In the third invention, in the second invention, the required torque is previously stored in the form of a map as a function of the depression amount of the accelerator pedal and the engine speed, and the calculating means calculates the required torque from this map. There is.

In the fourth invention, in the third invention,
The engine is equipped with a judging means for judging whether or not the engine is in an operating state where the engine load is substantially zero, and when the engine is judged to be in an operating state where the engine load is almost zero, the required torque calculated from the map becomes zero. On the other hand, when it deviates to the positive side, the value obtained by subtracting the deviation amount of the required torque from zero from the required torque calculated from the map is set as the actual required torque, and it is determined that the engine load is in an operating state of almost zero. When the required torque calculated from the map deviates to the negative side with respect to zero, the value obtained by adding the deviation amount of the required torque with respect to zero to the required torque calculated from the map is set as the actual required torque. There is.

In the fifth invention, in the first invention,
The fuel injection timing is calculated based on the calculated required torque and engine speed regardless of whether the first combustion or the second combustion is performed. In a sixth aspect of the invention, in the first aspect of the invention, the fuel injection pressure is changed based on the calculated required torque and engine speed regardless of which of the first combustion and the second combustion is being performed. It is calculated.

In the seventh invention, in the first invention,
A throttle valve is arranged in the engine intake passage, and the throttle valve is opened based on the calculated required torque and engine speed regardless of whether the first combustion or the second combustion is performed. The degree is calculated. In an eighth aspect based on the first aspect, an exhaust gas recirculation device for recirculating exhaust gas discharged from the combustion chamber into the engine intake passage is provided, and the inert gas is recirculated exhaust gas.

In the ninth invention, in the eighth invention,
The exhaust gas recirculation rate in the first combustion state is approximately 55% or more. According to a tenth aspect of the invention, in the eighth aspect, the exhaust gas recirculation device includes an exhaust gas recirculation control valve that controls the amount of exhaust gas recirculation, and the combustion of either the first combustion or the second combustion is performed. The opening degree of the exhaust gas recirculation control valve is calculated based on the calculated required torque and engine speed even when it is being performed.

According to the eleventh invention, in the first invention, a catalyst having an oxidizing function is arranged 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 the thirteenth invention, in the first invention, the operating region of the engine is divided into a first operating region on the low load side and a second operating region on the high load side, and the first combustion is performed in the first operating region. The second combustion is performed in the second operation area.

[0021]

FIG. 1 shows the 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, and 9 is an intake port. Indicates an exhaust valve, and 10 indicates an exhaust port, respectively. The intake port 8 is connected to a surge tank 12 via a corresponding intake branch pipe 11, and the surge tank 12 is connected via an intake duct 13 and an intercooler 14 to a supercharger, for example, an outlet portion of a compressor 16 of an exhaust turbocharger 15. Be connected. The 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 pipe 22 through the exhaust turbocharger 15
Of the exhaust turbine 23, and the outlet of the exhaust turbine 23 is connected via an exhaust pipe 24 to a catalytic converter 26 containing a catalyst 25 having an oxidizing function. 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 intake pipe 17 downstream of the throttle valve 20 are connected to each other via an EGR passage 29, and the EGR passage 29 is driven by a step motor 30. EGR
A control valve 31 is arranged. An intercooler 32 for cooling the EGR gas flowing in the EGR passage 29 is arranged in the EGR passage 29. In the embodiment shown in FIG. 1, engine cooling water is introduced into the intercooler 32,
The EGR gas is cooled by the engine cooling water.

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 electrically controlled variable discharge fuel pump 35, and the fuel supplied into the common rail 34 is supplied to the fuel injection valve 6 via each fuel supply pipe 33. A fuel pressure sensor 36 for detecting a fuel pressure in the common rail 34 is attached to the common rail 34, and a fuel pump 35 is arranged so that the fuel pressure in the common rail 34 becomes a target fuel pressure based on an output signal of the fuel pressure sensor 36. Is controlled.

The crankshaft 37 is connected to an automatic transmission 38 as shown in FIG. The electronic control unit 40 is composed of a digital computer, and has a ROM (read only memory) 42 and a RAM (random access memory) 4 connected to each other by a bidirectional bus 41.
3, CPU (microprocessor) 44, input port 4
5 and output port 46. Air-fuel ratio sensor 27
Is output 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 45 via the corresponding AD converter 47. . Also,
The input port 45 has a crank angle sensor 52 that generates an output pulse each time the crankshaft rotates, for example, 30 °.
Are connected. Further, the vehicle speed sensor 53 is connected to the input port 45.
Output pulse representing the vehicle speed of the
The output signal of the neutral sensor 54 for detecting whether or not 8 is in the neutral position is input to the input port 45. Further, the output signal of the idle switch 55 for detecting whether or not the throttle valve 20 is at the idling opening is also input to the input port 45. On the other hand, the output port 46 is connected to the fuel injection valve 6, the throttle valve control step motor 19, the EGR control valve control step motor 30 and the fuel pump 35 via the corresponding drive circuit 48.

FIG. 2 shows the throttle valve 2 during engine low load operation.
The change in output torque when the air-fuel ratio A / F (horizontal axis in FIG. 2) is changed by changing the opening degree of 0 and the EGR rate, and the change in the amount of smoke, HC, CO, and NO _x emissions. It represents the experimental example shown. As can be seen from FIG. 2, in this experimental example, the EGR rate becomes larger as the air-fuel ratio A / F becomes smaller, and when the air-fuel ratio is equal to or less than the theoretical air-fuel ratio (≈14.6), the EGR rate becomes 65% or more.

As shown in FIG. 2, when the air-fuel ratio A / F is reduced by increasing the EGR rate, the EGR rate becomes around 40%, and when the air-fuel ratio A / F reaches about 30, smoke is generated. The amount of generation begins to increase. Next, when the EGR rate is further increased and the air-fuel ratio A / F is made smaller, 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 sharply decreases this time, the EGR rate is increased to 65% or more, and the smoke becomes almost zero when the air-fuel ratio A / F is around 15.0. . That is, soot is hardly generated. At this time, the output torque of the engine slightly decreases, and N
Generation amount of O _X is considerably lower. On the other hand, at this time, HC,
The amount of CO generated starts to increase.

FIG. 3 (A) shows the change in combustion pressure in the combustion chamber 5 when the air-fuel ratio A / F is around 21 and the amount of smoke is largest, and FIG. 3 (B) shows the air-fuel ratio A / F. It shows a change in the combustion pressure in the combustion chamber 5 when F is around 18 and the amount of smoke generated is almost zero. As can be seen by comparing FIG. 3 (A) and FIG. 3 (B), in the case shown in FIG. 3 (B) where the amount of smoke generated is almost zero, the amount of smoke generated is large.
It can be seen that the combustion pressure is lower than in the case shown in (A).

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

Secondly, when the amount of smoke produced, that is, the amount of soot produced, becomes almost zero, as shown in FIG.
Emissions will increase. This means that hydrocarbons are discharged without growing to soot. That is, linear hydrocarbons and aromatic hydrocarbons contained in the fuel as shown in FIG. 4 are thermally decomposed to form soot precursors when the temperature is raised in a state of oxygen deficiency, and then mainly soot is formed. Soot consisting of a solid with carbon atoms gathered 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. After that, it will grow to soot. Therefore, as described above, when the amount of soot generated becomes almost zero, the emission amounts of HC and CO increase as shown in FIG. 2. At this time, HC is a soot precursor or a hydrocarbon in the state before it. .

Summarizing these considerations based on the experimental results shown in FIGS. 2 and 3, when the combustion temperature in the combustion chamber 5 is low, the soot generation amount becomes almost zero, and at this time, the soot precursor or the soot precursor The hydrocarbons in this state are discharged from the combustion chamber 5. As a result of further detailed experimental research on this, when the temperature of the fuel and the gas around it in the combustion chamber 5 is below a certain temperature, the soot growth process stops halfway, that is, the soot is generated. It was found that soot was not generated at all and soot was generated when the temperature of the fuel and its surroundings in the combustion chamber 5 reached a certain temperature or higher.

By the way, the temperature of the fuel and its surroundings when the hydrocarbon production process 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, the air-fuel ratio and the compression ratio. It cannot be said how many times it changes, but this certain temperature is closely related to the amount of NO _x produced, and therefore this certain temperature is defined to some extent from the amount of NO _x produced. be able to. That is, as the EGR rate increases, the temperature of the fuel during combustion and the gas surrounding it decrease, and the amount of NO _x generated decreases. At this time, soot is hardly generated when the amount of NO _x generated is about 10 p.pm or less. Therefore, the above certain temperature is NO
It is almost the same as the temperature when the amount of _X generation is around 10 p.pm or less.

Once soot is produced, this soot cannot be purified by a post-treatment using a catalyst having an oxidizing function. On the other hand, the soot precursor or the hydrocarbon in the state before it can be easily purified by a post-treatment using a catalyst having an oxidizing function. Considering the post-treatment with a catalyst having an oxidizing function as described above, it is extremely difficult to determine whether the hydrocarbon is discharged from the combustion chamber 5 in the state of the soot precursor or in the state before it, or is discharged from the combustion chamber 5 in the form of soot. There is a big difference. The new combustion system employed in the present invention allows hydrocarbons to be discharged from the combustion chamber 5 in the form of soot precursors or pre-presence conditions without producing soot in the combustion chamber 5 The core is to oxidize with a catalyst having an oxidizing function.

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

That is, if only air exists 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 locally becomes extremely high. That is, at this time, the air separated from the fuel hardly absorbs the combustion heat of the fuel. In this case, since the combustion temperature locally becomes extremely high, the unburned hydrocarbons that have received this heat of combustion generate soot.

On the other hand, the situation is slightly different when the fuel is present in a mixed gas of a large amount of inert gas and a small amount of air.
In this case, the evaporated fuel diffuses into the surroundings, reacts with oxygen mixed in the inert gas, and burns. In this case, the combustion heat is absorbed by the surrounding inert gas, so that the combustion temperature does not rise so much. 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 suppressed low by the endothermic action of the inert gas.

In this case, in order to suppress the temperature of the fuel and the gas around it to a temperature lower than the temperature at which soot is produced, an amount of inert gas sufficient to absorb the amount of heat required to do so is required. . Therefore, if the fuel amount increases, the required amount of inert gas also increases accordingly. In this case, the larger the specific heat of the inert gas, the stronger the endothermic action. Therefore, the inert gas is preferably a gas having a large specific heat. In this respect, since CO ₂ and EGR gas have relatively large specific heat, it can be said that it is preferable to use EGR gas as the inert gas.

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

As shown by the curve A in FIG. 5, when the EGR gas is strongly cooled, the soot generation amount reaches a peak when the EGR rate is slightly lower than 50%, and in this case, the EGR rate is almost 55. Almost no soot is generated if the percentage is exceeded. 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 when the EGR rate is slightly higher than 50%. In this case, the EGR rate is approximately 65% or more. If so, soot is hardly generated.

Further, as shown by the curve C in FIG. 5, EG
When the R gas is not forcibly cooled, the EGR rate is 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. Note that FIG. 5 shows the amount of smoke generated when the engine load is relatively high, and the EGR rate at which the amount of soot generated peaks when the engine load decreases and the EGR rate at which soot almost does not occur decreases. The lower limit of is also slightly lowered. Thus, the lower limit of the EGR rate at which soot is hardly generated changes depending on the cooling degree of EGR gas and the engine load.

FIG. 6 shows a mixture of EGR gas and air required to bring the temperature of the fuel and its surrounding gas at the time of combustion to a temperature lower than the temperature at which soot is generated when EGR gas is used as the inert gas. The amount of gas, the ratio of air in this mixed gas amount, and the ratio of EGR gas in this mixed gas are shown. Note that, in FIG. 6, the vertical axis represents the total intake gas amount sucked into the combustion chamber 5, and the chain line Y represents the total intake gas amount that can be sucked into the combustion chamber 5 when supercharging is not performed. ing. The horizontal axis shows the required load.

Referring to FIG. 6, the ratio of air, that is, the amount of air in the mixed gas indicates the amount of air required 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 theoretical air-fuel ratio. On the other hand, in FIG. 6, the ratio of EGR gas, that is, the amount of EGR gas in the mixed gas is set so that when the injected fuel is burned, the temperature of the fuel and its surrounding gas is lower than the temperature at which soot is formed. The minimum required EGR gas amount is shown. This EGR gas amount is approximately 55% or more when expressed by the EGR rate, and is 7 in the embodiment shown in FIG.
It is 0% or more. That is, the total intake gas amount sucked into the combustion chamber 5 is shown by a solid line X in FIG. 6, and the ratio of the air amount and the EGR gas amount in the total intake gas amount X is shown in FIG.
When the ratio is as shown in (1), the temperature of the fuel and the gas around it becomes lower than the temperature at which soot is generated, and thus soot is not generated at all. Further, the amount of NO _x generated at this time is around 10 p.pm or less, so N
Generation of O _X becomes extremely small.

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

By the way, when supercharging is not performed,
The upper limit of the total intake gas amount X drawn into the baking chamber 5 is Y.
Therefore, the required load is L in FIG.₀Greater than
In the region, the EGR gas ratio is increased as the required load increases.
It is possible to maintain the air-fuel ratio at the theoretical air-fuel ratio unless it is decreased.
Can not. In other words, it is necessary when supercharging is not done.
Load demand is L₀The theoretical air-fuel ratio in a region larger than
If you try to maintain the fuel ratio, the required load will increase.
As a result, the EGR rate decreases, so the required load is L ₀Than
In a large area, the temperature of the fuel and the gas around it is sooted
It cannot be maintained below the temperature at which it is produced.

However, as shown in FIG. 1, when EGR gas is recirculated to the inlet side of the supercharger, that is, in the air suction pipe 17 of the exhaust turbocharger 15 via the EGR passage 29, the required load is larger than L _0. EGR rate is 5 in the region
It can be maintained above 5 percent, for example 70 percent, thus maintaining the fuel and surrounding gas temperatures 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 intake gas boosted by the compressor 16 of the exhaust turbocharger 15 also becomes 70%. The temperature of the fuel and the gas around it can be maintained below the temperature at which soot is produced, up to the limit that can be boosted by the compressor 16. Therefore, the operating range of the engine capable of producing the low temperature combustion can be expanded.

In this case, when the EGR rate is set to 55% or more in the region where the required load is larger than L ₀ , the EGR control valve 31 is fully opened and the throttle valve 20
Is closed a little. 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 made smaller than the air amount shown in FIG. The generation amount of NO _X can be reduced to about 10 p.pm or less while preventing the generation of air, and the air amount can be increased more than the air amount shown in FIG. Even if lean from 17 to 18, the amount of NO _x generated can be reduced to about 10 p.pm 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 excessive fuel does not grow to soot, and soot is generated. There is no. Further, at this time, only a very small amount of NO _X is 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 becomes high, but in the present invention the combustion temperature is suppressed to a low temperature, soot Not generated at all. Furthermore, NO _X
Also produces only a very small amount.

Thus, 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. However, the amount of NO _x generated is 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 temperature of the gas around it can be suppressed below the temperature at which the growth of hydrocarbons stops halfway only when the engine is operating at low load, where the calorific value of combustion is relatively small. To be Therefore, in the embodiment according to the present invention, when the engine is operated at low load, the temperature of the fuel and the gas around it during combustion is suppressed below the temperature at which the growth of hydrocarbons stops halfway, and the first combustion, that is, low temperature combustion is performed. In addition, the second combustion, that is, the combustion normally performed from the conventional one, is performed during the engine high load operation. It should be noted that here, the first combustion, that is, low temperature combustion, as is clear from the above description, the amount of inert gas in the combustion chamber is larger than the amount of inert gas at which the amount of soot generated peaks, and soot is almost generated. The second combustion, that is, the combustion that is normally performed in the past, is the combustion that does not have the amount of soot generated and the amount of the inert gas in the combustion chamber is smaller than the amount of the inert gas that does not reach the peak. Say that. FIG. 7 shows a first operating region I in which the first combustion, that is, low temperature combustion is performed, and a second operating region II in which the second combustion, that is, combustion by the conventional combustion method is performed.
Is shown. In FIG. 7, the vertical axis TQ shows the required torque, and the horizontal axis N shows the engine speed.
Further, in FIG. 7, X (N) is the first operating region I and the second operating region I.
The second boundary between the first operating area I and the second operating area II is indicated by Y (N). The change determination of the operating region from the first operating region I to the second operating region II is performed based on the first boundary X (N), and the change from the second operating region II to the first operating region I is performed. The change judgment of the operating region is made based on the second boundary Y (N).

That is, the operating condition of the engine is the first operating region I.
Required torque TQ when low temperature combustion is performed
Is above the first boundary X (N) which is a function of the engine speed N, it is judged that the operating region has moved to the second operating region II,
Combustion is performed by a conventional combustion method. Next, the required torque TQ is a second boundary Y (N) which is a function of the engine speed N.
When it becomes lower than that, it is determined that the operating region has moved to the first operating region I, and low temperature combustion is performed again.

In this way, the two boundaries of the first boundary X (N) and the second boundary Y (N) on the lower torque side of the first boundary X (N) are provided as follows. For one reason. The first reason is that the combustion temperature is relatively high on the high torque side of the second operating region II, and at this time the required torque TQ is the first boundary X.
This is because even if the temperature becomes lower than (N), low temperature combustion cannot be performed immediately. That is, the low temperature combustion is not started immediately unless the required torque TQ becomes considerably low, that is, when it becomes lower than the second boundary Y (N). The second reason is the first operating region I and the second operating region II.
This is because a hysteresis is provided for changes in the operating region between.

By the way, when the engine is operating in the first operating region I and low-temperature combustion is performed, soot is hardly generated, and instead, unburned hydrocarbons are in the state of the soot precursor or the state before it. It is discharged from the combustion chamber 5 in shape. At this time, the unburned hydrocarbons discharged from the combustion chamber 5 are satisfactorily oxidized by the catalyst 25 having an oxidizing function. Catalyst 2
The 5 it is possible to use an oxide catalyst, three-way catalyst, or _the NO _X absorbent. _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
It has a function of releasing NO _x when the average air-fuel ratio in the inside becomes rich.

This NO _x absorbent uses, for example, alumina as a carrier, and potassium K, sodium N, etc. are supported on this carrier.
a, at least one selected from alkali metals such as lithium Li and 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.

Not only oxidation catalysts, but also three-way catalysts and NO
_{The X} absorbent also has an oxidizing function, so that the three-way catalyst and the 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 opening of the throttle valve 20, the opening of the EGR control valve 31, the EGR rate, the air-fuel ratio, the injection timing and the injection amount with respect to the required torque TQ. As shown in FIG. 9, in the first operating region I where the required torque TQ is low, the opening degree of the throttle valve 20 is gradually increased from near full closing to about 2/3 opening degree as the required load L increases. The opening degree of the EGR control valve 31 is gradually increased from near full close to full open as the required load L increases.
Further, in the example shown in FIG. 9, EG in the first operating region I
The R ratio is set to about 70%, and the air-fuel ratio is a slightly lean lean air-fuel ratio.

In other words, in the first operating region I, EGR
The opening of the throttle valve 20 and the opening of the EGR control valve 31 are controlled so that the ratio 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 degree of the EGR control valve 31 based on the output signal of the air-fuel ratio sensor 27. Further, in the first operation region I, fuel injection is performed before the compression top dead center TDC. In this case, the injection start timing θS becomes late as the required load L becomes high, and the injection completion timing θE also becomes late as the injection start timing θS becomes late.

During the idling operation, the throttle valve 20 is closed to the fully closed state, and at this time, the EGR control valve 31 is also closed to the fully closed state. Throttle valve 2
When 0 is closed to near full closure, the pressure in the combustion chamber 5 at the beginning of compression becomes low and the compression pressure becomes small. When the compression pressure becomes small, the compression work by the piston 4 becomes small, so that the vibration of the engine body 1 becomes small. That is, during idling operation, the throttle valve 20 is closed close to the fully closed state in order to suppress the vibration of the engine body 1.

On the other hand, the operating region of the engine is the first operating region I.
When changing from the second operating range II to the second operating range II, the opening degree of the throttle valve 20 is increased stepwise from about 2/3 opening degree 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, and the air-fuel ratio is increased stepwise. That is, the EGR that produces a large amount of smoke with an EGR rate
The operating range of the engine is the first because the rate range (Fig. 5) is skipped.
A large amount of smoke does not occur when changing from the operating region I to the second operating region II.

In the second operation region II, the second combustion, that is, the combustion which is conventionally performed is performed. Although soot and NO _X are slightly generated in this combustion method, the thermal efficiency is higher than that in low temperature combustion, and therefore, when the operating region of the engine changes from the first operating region I to the second operating region II, as shown in FIG. In addition, the injection amount is reduced stepwise. In the second operating region II, the throttle valve 20 is kept fully open except for a part, and the opening degree of the EGR control valve 31 is gradually reduced as the required torque TQ increases. Also, this operating area II
Then, the EGR rate becomes lower as the required torque TQ becomes higher, and the air-fuel ratio becomes smaller as the required torque TQ becomes higher. However, the air-fuel ratio is set to the lean air-fuel ratio even if the required torque TQ becomes high. Further, in the second operation region II, the injection start timing θS is set near the compression top dead center TDC.

FIG. 10A shows the relationship between the required torque TQ, the depression amount L of the accelerator pedal 50, and the engine speed N. In FIG. 10A, each curve represents an equal torque curve, the curve indicated by TQ = 0 indicates that the torque is zero, and the remaining curves indicate TQ =.
The required torque gradually increases in the order of a, TQ = b, TQ = c, TQ = d. Required torque T shown in FIG. 10 (A)
Q is the accelerator pedal 50 as shown in FIG.
It is stored in advance in the ROM 42 in the form of a map as a function of the depression amount L and the engine speed N. In the present invention, FIG.
The required torque TQ corresponding to the depression amount L of the accelerator pedal 50 and the engine speed N is first calculated from the map shown in (B), and the fuel injection amount and the like are calculated based on this required torque TQ.

FIG. 11 shows the air-fuel ratio A / F in the first operating region I. In FIG. 11, A / F = 15.
5, each curve shown by A / F = 16, A / F = 17, A / F = 18 has an air-fuel ratio of 15.5, 16, 17, 18 respectively.
And the air-fuel ratio between the curves is determined by proportional distribution. As shown in FIG. 11, 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 is lean.

That is, the lower the required torque TQ, the smaller the amount of heat generated by combustion. Therefore, as the required torque TQ becomes lower, the low temperature combustion can be performed even if the EGR rate is lowered. If the EGR rate is reduced, the air-fuel ratio will increase,
Therefore, as shown in FIG. 11, the air-fuel ratio A / F is increased as the required torque TQ decreases. Air-fuel ratio A / F
Becomes larger, the fuel consumption rate increases. Therefore, in order to make the air-fuel ratio as lean as possible, in the embodiment of the present invention, the air-fuel ratio A / F is made larger as the required torque TQ becomes lower.

[0063] Figure 12 (A) shows the injection amount Q in the first operating region I, and FIG. 12 (B) is shows the injection start timing θS in the first operating region I. 12
As shown in (A), the injection amount Q in the first operating region I is stored in advance in the ROM 42 in the form of a map in the form of a map as a function of the required torque TQ and the engine speed N.
As shown in FIG. 2 (B), the injection start timing θS in the first operating region I is also stored in advance in the ROM 42 in the form of a map as a function of the required torque TQ and the engine speed N.

Further, the target opening degree ST of the throttle valve 20 required to bring the air-fuel ratio to the target air-fuel ratio shown in FIG.
As shown in FIG. 3 (A), it is stored in the ROM 42 in advance in the form of a map as a function of the required torque TQ and the engine speed N, and the EGR required to bring the air-fuel ratio to the target air-fuel ratio shown in FIG. The target opening degree SE of the control valve 31 is shown in FIG.
As shown in (B), it is stored in advance in the ROM 42 in the form of a map as a function of the required torque TQ and the engine speed N.

Further, in the embodiment according to the present invention, the fuel injection pressure in the first operating region I, that is, the target fuel pressure P in the common rail 34, is the required torque TQ and the engine speed N as shown in FIG. 13 (C). It is stored in advance in the ROM 42 in the form of a map as a function. FIG. 14 shows the second combustion,
That is, it shows the target air-fuel ratio when normal combustion is performed by the conventional combustion method. In addition, in FIG.
= 24, A / F = 35, A / F = 45, and A / F = 60, the respective curves indicated by the target air-fuel ratios of 24, 35, 45, and 6 respectively.
0 is shown.

FIG. 15A shows the injection amount Q in the second operating region II, and FIG. 15B shows the second operating region.
The injection start timing θS in II is shown . Figure 15
As shown in (A), the injection amount Q in the second operation region II is stored in advance in the ROM 42 in the form of a map as a function of the required torque TQ and the engine speed N, and FIG.
As shown in FIG. 5 (B), the injection start timing θS in the second operating region II is also stored in advance in the ROM 42 in the form of a map as a function of the required torque TQ and the engine speed N.

Further, the target opening degree ST of the throttle valve 20 required to bring the air-fuel ratio to the target air-fuel ratio shown in FIG.
As shown in FIG. 6 (A), it is stored in the ROM 42 in advance in the form of a map as a function of the required torque TQ and the engine speed N, and the EGR required to make the air-fuel ratio the target air-fuel ratio shown in FIG. The target opening degree SE of the control valve 31 is shown in FIG.
As shown in (B), it is stored in advance in the ROM 42 in the form of a map as a function of the required torque TQ and the engine speed N.

Further, in the embodiment according to the present invention, the fuel injection pressure in the second operating region II, that is, the target fuel pressure P in the common rail 34, is the required torque TQ and the engine speed N as shown in FIG. 16 (C). It is stored in advance in the ROM 42 in the form of a map as a function. In the embodiment according to the present invention, the injection amount Q is calculated from the map shown in FIG. 12 (A) in the first operating region I, and the injection amount Q from the map shown in FIG. 15 (A) in the second operating region II. Is calculated. The maps shown in both FIG. 12 (A) and FIG. 15 (A) are a function of the required torque TQ and the engine speed N, and therefore, at the first boundary X (N) or the second boundary Y (N). 1st combustion to 2nd combustion or 2nd combustion to 1st
Even if the combustion is switched to combustion, the required torque does not change at the time of switching, and thus there is no risk of torque fluctuation at the time of switching.

The injection amount Q for the first combustion shown in FIG. 12A is also the injection amount Q for the second combustion shown in FIG. 15A.
Is also obtained by experiment. In this case, the injection amount Q for the first combustion and the injection amount Q for the second combustion are each at the boundary X.
Since the torque difference between (N) and Y (N) can be individually obtained by an experiment without any consideration, FIG.
The maps shown in FIGS. 15A and 15A can be created very easily.

On the other hand, the required torques indicated by the curves in FIG. 10A are empirically determined so as to ensure good drivability of the vehicle. Therefore, TQ = 0 in FIG.
The actual engine output torque may not be zero on the curve indicated by. Therefore, in the embodiment according to the present invention, it is determined whether or not the engine is in an operating state where the engine load is almost zero,
When the required torque TQ calculated from the map shown in FIG. 10 (A) is deviated to the positive side with respect to zero when it is determined that the engine is in an operating state where the engine load is almost zero, FIG.
A value obtained by subtracting the deviation amount of the required torque TQ from zero from the required torque TQ calculated from the map shown in (A) is set as the actual required torque, and when it is determined that the engine load is in an operating state of almost zero. When the required torque TQ calculated from the map shown in FIG. 10 (A) deviates to the negative side with respect to zero, the required torque TQ calculated from the map shown in FIG. 10 (A) deviates from the required torque TQ with respect to zero. The value obtained by adding the quantities is used as the actual required torque.

More specifically, for example, in FIG. 10 (A), when the actual output torque of the engine is zero, the accelerator pedal 50 is depressed by L ₀ and the engine speed is indicated by F at N _0. If it is in the operating state, at this time, FIG.
The required torque TQ on the map shown in (B) does not become zero, and the required torque TQ on the map deviates toward the positive side by ΔKG with respect to zero. In this case, in the embodiment according to the present invention, a value (TQ-ΔKG) obtained by subtracting the required torque deviation amount ΔKG from the required torque TQ calculated from the map shown in FIG. 10B is set as the actual required torque. As a result, it is possible to obtain the originally intended engine output torque with respect to the depression amount L of each accelerator pedal 50 and each engine speed N.

Next, the operation control will be described with reference to FIG. Referring to FIG. 17, first step 1
At 00, it is determined whether or not the 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
In step 1, it is determined whether the required load L has become larger than the first boundary X (N). When L ≦ X (N), the routine proceeds to step 103, where low temperature combustion is performed.

That is, at step 103, the required torque TQ is calculated from the map shown in FIG. Next, at step 104, the required torque TQ calculated from the map.
Value (= TQ-
ΔKG) is the final required torque TQ. Next, in steps 105 to 109, the final required torque is used to calculate the throttle valve opening ST, the EGR control valve opening SE, the injection amount Q, the injection start timing θS, and the fuel injection pressure P.

That is, at step 105, the target opening degree ST of the throttle valve 20 is calculated from the map shown in FIG. 13 (A), and the opening degree of the throttle valve 20 is made this target opening degree ST. Next, at step 106, the target opening degree SE of the EGR control valve 31 is calculated from the map shown in FIG.
The opening degree of the EGR control valve 31 is set to this target opening degree SE.
Next, at step 107, the injection amount Q is calculated from the map shown in FIG. Next, at step 108, the injection start timing θS is calculated from the map shown in FIG. Next, at step 109, the target fuel pressure in the common rail 34, that is, the injection pressure P is calculated from the map shown in FIG. 13 (C).

Next, at step 110, it is judged from the output signal of the idle switch 55 whether or not the throttle valve 20 is at the idling opening degree. When the throttle valve 20 is at the idling opening, the routine proceeds to step 111, where it is judged from the output pulse of the vehicle speed sensor 53 whether the vehicle speed is zero. When the vehicle speed is zero, the routine proceeds to step 112, where it is judged from the output signal of the neutral switch 54 whether or not the automatic transmission 38 is in the neutral position.
When the automatic transmission 38 is in the neutral position, the routine proceeds to step 113, where it is determined whether or not a fixed time has elapsed since the automatic transmission 38 was in the neutral position. When the fixed time has elapsed, the engine load is substantially zero. It is determined that the vehicle is operating.

When it is determined that the engine is operating under almost zero engine load, the routine proceeds to step 114, and FIG.
The required torque TQ is calculated from the map shown in (B), and then, at step 115, the required torque TQ is deviated by a deviation amount Δ.
KG. On the other hand, in step 101, L> X
If it is determined that (N) is reached, the routine proceeds to step 102, where the flag I is reset, and then step 118
Then, the second combustion is performed.

That is, at step 118, the required torque TQ is calculated from the map shown in FIG. 10 (B). Next, at step 119, the required torque TQ calculated from the map.
Value (= TQ-
ΔKG) is the final required torque TQ. Next, in steps 120 to 124, the final required torque is used to open the throttle valve opening ST, the EGR control valve body opening SE, the injection amount Q, the injection start timing θS, and the fuel injection pressure P.
Is calculated.

That is, at step 120, the target opening degree ST of the throttle valve 20 is calculated from the map shown in FIG. 16 (A), and the opening degree of the throttle valve 20 is made this target opening degree ST. Next, at step 121, the target opening degree SE of the EGR control valve 31 is calculated from the map shown in FIG. 16 (B),
The opening degree of the EGR control valve 31 is set to this target opening degree SE.
Next, at step 122, the injection amount Q is calculated from the map shown in FIG. Next, at step 123, the injection start timing θS is calculated from the map shown in FIG. 15 (B). Next, at step 124, the target fuel pressure in the common rail 34, that is, the injection pressure P is calculated from the map shown in FIG. 16 (C).

When the flag I is reset, in the next processing cycle, the routine proceeds from step 100 to step 116, where it is judged if the required load L has become lower than the second boundary Y (N). When L ≧ Y (N), step 118
And the second combustion is performed under a lean air-fuel ratio.
On the other hand, when it is determined at step 116 that L <Y (N), the routine proceeds to step 117, where the flag I is set, then the routine proceeds to step 103, where low temperature combustion is performed.

[0080]

The output torque can be prevented from fluctuating at the time of switching between the first combustion and the second combustion without requiring complicated work.

[Brief description of drawings]

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

FIG. 2 is a diagram showing an amount of smoke and NO _X generated, and the like.

FIG. 3 is a diagram showing a combustion pressure.

FIG. 4 is a diagram showing a fuel molecule.

FIG. 5 is a diagram showing a relationship between a smoke generation amount 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 operating region I and a second operating 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 a required torque.

11 is a diagram showing an air-fuel ratio in a first operating region I. FIG.

FIG. 12 is a diagram showing a map of an injection amount and the like.

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

FIG. 14 is a diagram showing an air-fuel ratio in the second combustion.

FIG. 15 is a diagram showing a map of an injection amount and the like.

FIG. 16 is a diagram showing a map of a target opening degree etc. of a throttle valve.

FIG. 17 is a flowchart for controlling the operation of the engine.

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

FIG. 19 is a diagram showing a map of an injection amount.

[Explanation of symbols]

6 ... Fuel injection valve 20 ... Throttle valve 29 ... EGR passage

─────────────────────────────────────────────────── ─── Continuation of front page (51) Int.Cl. ⁷ Identification code FI F01N 3/24 F01N 3/24 S F02D 21/08 301 F02D 21/08 301D 41/02 380 41/02 380E 41/40 41 / 40 F 43/00 301 43/00 301H 301N F02M 25/07 570 F02M 25/07 570D (72) Inventor Takekazu Ito 1 Toyota-cho, Toyota-shi, Aichi Toyota Motor Corporation (72) Inventor Hiroki Murata 1 Toyota-cho, Toyota-shi, Aichi Toyota Motor Co., Ltd. (56) Reference JP-A-7-4287 (JP, A) JP-A-8-177654 (JP, A) JP-A-8-86251 (JP, A) JP 9-287527 (JP, A) JP 9-287528 (JP, A) (58) Fields investigated (Int.Cl. ⁷ , DB name) F02B 1/00-23/06 F02D 41 / 00-45/00 F02M 25/07 570

Claims (13)

(57) [Claims] 1. When the amount of inert gas in the combustion chamber increases, the soot generation amount gradually increases and reaches a peak, and when the amount of inert gas in the combustion chamber further increases, combustion in the combustion chamber occurs. In an internal combustion engine in which the temperature of the fuel and the gas around it are lower than the soot generation temperature and soot is hardly generated, the amount of inert gas in the combustion chamber is higher than the amount of inert gas at which the soot generation peaks. A switching means for selectively switching between the first combustion in which much 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 soot generated reaches a peak, and an engine. And a calculation means for calculating the required torque of the engine, and the fuel injection amount is calculated based on the calculated required torque and the engine speed regardless of whether the first combustion or the second combustion is performed. Internal combustion engine Seki. 2. The internal combustion engine according to claim 1, wherein the calculating means calculates the required torque based on the depression amount of the accelerator pedal and the engine speed. 3. The internal combustion engine according to claim 2, wherein the required torque is stored in advance in the form of a map as a function of the depression amount of the accelerator pedal and the engine speed, and the calculating means calculates the required torque from the map. . 4. A determination means for determining whether or not the engine load is in an operating state where the engine load is substantially zero, and is calculated from the map when it is determined that the engine load is in an operating state where the engine load is substantially zero. When the required torque deviates to the positive side with respect to zero, the value obtained by subtracting the deviation amount of the required torque from zero from the required torque calculated from the map is taken as the actual required torque, and the engine load is almost zero. When it is determined that the required torque calculated from the map deviates to the negative side with respect to zero, the actual torque is calculated by adding the amount of deviation of the required torque with respect to zero to the required torque calculated from the map. The internal combustion engine according to claim 3, wherein the required torque is obtained. 5. The fuel injection timing is calculated based on the calculated required torque and engine speed regardless of whether the first combustion or the second combustion is being performed. Internal combustion engine according to. 6. The fuel injection pressure is calculated on the basis of the calculated required torque and the engine speed regardless of which of the first combustion and the second combustion is being performed. Internal combustion engine according to. 7. A throttle valve is arranged in the engine intake passage, and is based on the calculated required torque and engine speed regardless of whether the first combustion or the second combustion is being performed. The internal combustion engine according to claim 1, wherein the opening degree of the throttle valve is calculated by the above. 8. An exhaust gas recirculation device for recirculating exhaust gas discharged from a combustion chamber into an engine intake passage,
The internal combustion engine according to claim 1, wherein the inert gas comprises recirculated exhaust gas. 9. The internal combustion engine according to claim 8, wherein the exhaust gas recirculation rate in the first combustion state is approximately 55% or more. 10. The exhaust gas recirculation device comprises an exhaust gas recirculation control valve for controlling the amount of exhaust gas recirculation,
9. The opening degree of the exhaust gas recirculation control valve is calculated based on the calculated required torque and the engine speed regardless of whether the combustion is being performed or the second combustion is performed. Internal combustion engine described. 11. The internal combustion engine according to claim 1, wherein a catalyst having an oxidizing function is arranged in the engine exhaust passage. 12. The catalyst is an oxidation catalyst, a three-way catalyst or NO.
The internal combustion engine of claim 11, comprising at least one _X absorbent. 13. The engine operating region is divided into a first operating region on the low load side and a second operating region on the high load side, and first combustion is performed in the first operating region to perform the second operating region. Second in the area
The internal combustion engine according to claim 1, wherein the combustion is performed.