JP2000064884A - Internal combustion engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To perform stable combustion without causing smoke and a misfire. SOLUTION: First combustion generating little soot with an EGR gas amount in a combustion chamber 5 larger than an EGR gas amount having a peak generation amount of soot and second combustion with an EGR gas amount in the combustion chamber 5 smaller than an EGR gas amount having a peak generation amount of soot are selectively performed. A cooling device 23 is arranged in an EGR passage 21, and a cooling capacity necessary for maintaining a temperature of EGR gas flowing out of the cooling device 23 to a predetermine set temperature is provided to the cooling device 23. Injection timing becoming optimum when the temperature of EGR gas flowing out of the cooling device 23 is this preset temperature is previously stored, to determine the injection timing based on the stored injection timing.

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 amount of NO _x generated decreases. Therefore, the higher the EGR rate, the lower the amount of NO _x generated.

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 stipulated that the amount of _x and smoke generated should be as small as possible. 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 is still left.
The O _x, and 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
It has also been found that the amount of O _x generated is extremely small.
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, the fuel injection timing at which stable combustion is obtained is limited to a relatively narrow crank angle range. That is, if the fuel injection timing is advanced, the injected fuel is heated by the compressed high temperature gas for a long time, so the temperature of the fuel and the gas around it becomes high during combustion, and as a result, hydrocarbons grow to soot. Smoke will occur. On the other hand, when the fuel injection timing is delayed, most of the fuel does not burn because the temperature of the injected fuel does not rise so much, which results in misfire.

That is, in this new combustion, there is an optimum crank angle range in which smoke is not generated with respect to the fuel injection timing and stable combustion can be obtained without causing misfire. Therefore, when performing this new combustion, Needs to perform fuel injection in this optimum crank angle range. However, the optimum crank angle range in which this stable combustion is obtained is the value of the operating parameter of the engine that affects the temperature of the fuel and the gas around it during combustion in the combustion chamber, such as the air-fuel ratio, the EGR rate, and the intake air It changes depending on the temperature, the EGR gas temperature, the cooling water temperature of the engine, the humidity in the intake air, and the like.

For example, the higher the intake air temperature, the higher the gas temperature in the combustion chamber, and the higher the temperature of the injected fuel. In this case, in order to prevent the smoke from being generated, it is necessary to shorten the heating time of the injected fuel by the gas in the combustion chamber as the intake air temperature becomes higher, and therefore, the fuel injection timing becomes higher as the intake air temperature becomes higher. I need to be late.

On the other hand, the lower the intake air temperature, the lower the gas temperature in the combustion chamber, and the lower the temperature of the injected fuel. In this case, in order to prevent misfire, it is necessary to lengthen the heating time of the injected fuel by the gas in the combustion chamber as the intake air temperature becomes lower, and therefore the fuel injection timing becomes earlier as the intake air temperature becomes lower. There is a need to.

As described above, the optimum crank angle range for fuel injection varies depending on the values of the operating parameters of the engine that affect the temperature of the fuel during combustion in the combustion chamber and the gas temperature around it. In this case, if the value of the operating parameter of the engine changes drastically, even if the temperature of the fuel during combustion in the combustion chamber and the temperature of the gas around it change, the injection timing can be changed by following the change due to detection delay and the like. Therefore, smoke is generated because the injection timing deviates from the optimum crank angle range, or misfire occurs.

By the way, the intake air temperature, the cooling water temperature of the engine, and the humidity of the intake air among the engine operating parameters that affect the temperature of the fuel and the gas around it in the combustion chamber do not change suddenly, Therefore, even if these parameters change, the injection timing can be controlled within the optimum crank angle range by following the changes. Also,
When new combustion is being performed, the air-fuel ratio and EGR rate do not change much even if the operating conditions change. Therefore, even if these parameters change, the injection timing is adjusted to the optimum crank angle range by following the changes. Can be controlled within.

However, among the operating parameters of the engine that affect the temperature of fuel and the gas around it during combustion in the combustion chamber, the EGR gas temperature changes drastically when the operating state of the engine changes. Therefore, even if the injection timing is controlled according to the change in the EGR gas temperature when the EGR gas temperature changes, the injection timing cannot be changed by following the change in the EGR gas temperature. There is a problem that smoke is generated or misfire occurs due to deviation from the crank angle range.

[0016]

In order to solve the above problems, the first invention comprises an exhaust gas recirculation device for recirculating exhaust gas discharged from the engine into the engine intake passage, When the amount of recirculated exhaust gas supplied to the combustion chamber increases, the amount of soot gradually increases and reaches a peak.In an internal combustion engine, the amount of soot generated reaches a peak. The amount of recirculated exhaust gas supplied to the engine is increased, and a cooling device is installed in the exhaust gas recirculation passage that connects the engine exhaust passage and the engine intake passage. The optimum injection timing is stored in advance when the temperature of exhaust gas that has flowed out of the cooling device has the set temperature and the cooling capacity required to maintain the temperature of the exhaust gas at the preset temperature is maintained. Been Ri, so that determine the injection timing based on the injection timing stored.

That is, even if the exhaust gas fluctuates drastically, the temperature of the recirculated exhaust gas recirculated in the combustion chamber is maintained at the set temperature, and the optimum injection timing when the temperature of the recirculated exhaust gas is the set temperature. The injection timing is stored in advance, and the injection timing is determined based on the stored injection timing. Therefore, the injection timing is always the optimum injection timing even if the exhaust gas temperature fluctuates drastically.

In the second invention, in the first invention,
The engine cooling water is guided to the cooling device, and the recirculated exhaust gas is cooled by the engine cooling water. In the third invention, in the second invention, the cooling water temperature in the engine body is maintained substantially constant, and is substantially equal to the cooling water temperature in the engine body in which the set temperature is maintained substantially constant.

In the fourth invention, in the first invention,
Equipped with a detection means for detecting the value of the operating parameter of the engine that affects the temperature of the fuel and the gas around it during combustion in the combustion chamber, the injection timing is based on the value of the operating parameter, the injection timing at which smoke is generated, and misfire. Is controlled during the injection timing. In the fifth invention, in the fourth invention, the operating parameter is the air-fuel ratio, and the injection timing is delayed as the air-fuel ratio increases.

In the sixth invention, in the fourth invention,
The operating parameter is the exhaust gas recirculation rate, and the higher the exhaust gas recirculation rate, the earlier the injection timing. In the seventh invention, in the fourth invention, the operating parameter is the intake air temperature, and the injection timing is delayed as the intake air temperature increases. 4th in the 8th invention
In the second invention, the operating parameter is the engine cooling water temperature, and the injection timing is delayed as the engine cooling water temperature increases.

In the ninth invention, in the fourth invention,
The operating parameter is the humidity of the intake air, and the injection start timing is advanced as the humidity of the intake air increases. According to a tenth invention, in the first invention, a switching means for switching the flow direction of the recirculated exhaust gas in the cooling device to the opposite direction is provided, and the recirculation in the cooling device is performed every time a predetermined period elapses. The flow direction of the circulating exhaust gas is switched to the opposite direction.

According to the eleventh invention, in the first invention, the exhaust gas recirculation rate is about 55% or more. In the twelfth aspect of the invention, in the first aspect of the invention, a catalyst having an oxidizing function is arranged in the engine exhaust passage. Thirteen
According to the twelfth invention, in the twelfth invention, the catalyst comprises at least one of an oxidation catalyst, a three-way catalyst and a NO _x absorbent.

According to a fourteenth invention, in the first invention, the recirculation exhaust gas amount supplied to the combustion chamber is larger than the recirculation exhaust gas amount at which the soot generation amount reaches a peak, and soot is hardly generated. A switching means is provided for selectively switching between combustion and second combustion in which the amount of recirculated exhaust gas supplied into the combustion chamber is smaller than the amount of recirculated gas at which the amount of soot generated peaks.

According to a fifteenth invention, in the fourteenth invention, 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 operating range is divided into the first operating range. Is performed, and the second combustion is performed in the second operation region.

[0025]

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 to an air cleaner 14 via an intake duct 13. A throttle valve 16 driven by a step motor 15 is arranged in the intake duct 13. On the other hand, the exhaust port 10 is connected to the exhaust manifold 17
Also, it is connected via an exhaust pipe 18 to a catalytic converter 20 containing a catalyst 19 having an oxidizing function.

The exhaust manifold 17 and the surge tank 12 are connected to each other via an EGR passage 21, and an electric control type EGR control valve 22 is arranged in the EGR passage 21.
A cooling device 23 for cooling the EGR gas is arranged in the EGR passage 21. In the embodiment shown in FIG. 1, engine cooling water is introduced into the cooling device 23 from the cooling water inlet 24, and the EGR gas is cooled by this cooling water. The cooling water that has cooled the EGR gas is discharged from the cooling water discharge port 25 and returned to the engine body 1.

On the other hand, the fuel injection valve 6 is connected to a fuel reservoir, a so-called common lane 27, through a fuel supply pipe 26. Fuel is supplied to the common lane 27 from an electrically controlled variable fuel discharge fuel pump 28, and the fuel supplied to the common lane 28 is supplied to the fuel injection valve 6 through each fuel supply pipe 26. . A fuel pressure sensor 29 for detecting the fuel pressure in the common lane 27 is attached to the common lane 27 so that the fuel pressure in the common lane 27 becomes the target fuel pressure based on the output signal of the fuel pressure sensor 29. The discharge amount of the fuel pump 28 is controlled.

The electronic control unit 40 is composed of a digital computer, and has a ROM (Read Only Memory) 42, a RAM (Random Access Memory) 43, a CPU (Microprocessor) 44, an input port 45, and an input port 45 which are connected to each other by a bidirectional bus 41. The output port 46 is provided. As shown in FIG. 1, the output signal of the fuel pressure sensor 29 is input to the input port 45 via the corresponding AD converter 47. A water temperature sensor 30 for detecting the engine cooling water temperature is arranged in the engine body 1, and the output signal of the water temperature sensor 30 is input to the input port 45 via the corresponding AD converter 47. A mass flow rate detector 31 for detecting the mass flow rate of the intake air, a temperature sensor 32 for detecting the intake air temperature, and a humidity of the intake air are provided in the intake duct 13 upstream of the throttle valve 16. A humidity sensor 33 is arranged, and the output signals of the mass flow rate detector 31, the temperature sensor 32, and the humidity sensor 33 are input to the input port 45 via the corresponding AD converters 47, respectively.

The accelerator pedal 50 includes an accelerator pedal 5
A load sensor 51 that generates an output voltage proportional to the stepping amount L of 0 is connected, and the output voltage of the load sensor 51 is input to the input port 45 via the corresponding AD converter 47. A crank angle sensor 52 that generates an output pulse each time the crankshaft rotates, for example, 30 ° is connected 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 15, the EGR control valve 22 and the fuel pump 28 via the corresponding drive circuit 48.

FIG. 2 shows the throttle valve 1 during engine low load operation.
6 changes in the output torque when the air-fuel ratio A / F (horizontal axis in FIG. 2) is changed by changing the opening degree and EGR rate, and changes 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 becomes 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
The amount of O _x generated is considerably low. 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 18 and the amount of smoke generated is the 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 13 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 generated O _x 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. .

When these considerations based on the experimental results shown in FIGS. 2 and 3 are summarized, the soot generation amount becomes almost zero when the combustion temperature in the combustion chamber 5 is low, 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 has a deep relationship with 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 around it decreases, 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 generated, 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 generated. 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, and 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.

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 the gas around it to a temperature lower than the temperature at which soot is produced 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. In addition, the horizontal axis represents the required load, and Z1 represents the low load operation region.

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
The amount of O _x generated is extremely small.

When the fuel injection amount increases, the amount of heat generated when the fuel burns also increases. Therefore, 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.

On the other hand, in the load region Z2 of FIG. 6, the total intake gas amount X required to prevent the generation of soot exceeds the total intake gas amount Y that can be inhaled. Therefore, in this case, in order to supply the total intake gas amount X required to prevent the generation of soot into the combustion chamber 5, both the EGR gas and the intake air, or EG
It is necessary to supercharge or pressurize the R gas. When the EGR gas or the like is not supercharged or pressurized, the total intake gas amount X matches the total intakeable gas amount Y in the load region Z2. Therefore, in this case, in order to prevent the generation of soot, the air amount is slightly decreased to increase the EGR gas amount and the fuel is burned under the rich air-fuel ratio.

As described above, FIG. 6 shows the case where the fuel is burned under the stoichiometric air-fuel ratio. However, in the low load operation region Z1 shown in FIG. 6, the air amount is smaller than the air amount shown in FIG. At least, that is, even if the air-fuel ratio is made rich, the generation amount of NO _x can be reduced to around 10 p.pm or less while preventing the generation of soot, and the air can be reduced in the low load region Z1 shown in FIG. Even if the amount is made larger than the air amount shown in FIG. 6, that is, even if the average value of the air-fuel ratio is lean from 17 to 18, the amount of NO _x generated is 10 while the generation of soot is prevented.
It can be around or below ppm.

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, 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.

As described above, in the engine low load operation region Z1, 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. Therefore, 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 and the gas around it during combustion in the combustion chamber can be suppressed below the temperature at which the growth of hydrocarbons stops halfway only when the amount of heat generated by combustion is small and the engine load is relatively low. Therefore, in the present invention, when the engine load is relatively low, 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, When the engine load is relatively high, the second combustion, that is, the combustion that is more commonly performed than before is performed. 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 usually performed conventionally, is the combustion in which the amount of inert gas in the combustion chamber is smaller than the amount of inert gas in which the soot generation peaks. Say.

FIG. 7 shows a first operating region I in which the first combustion, that is, low temperature combustion is performed, and a second combustion region II in which the second combustion, that is, combustion by the conventional combustion method is performed. There is. In FIG. 7, the vertical axis L represents the depression amount of the accelerator pedal 50, that is, the required load, and the horizontal axis N represents the engine speed. Further, in FIG. 7, X (N) is the first
Shows the first boundary between the operating region I and the second operating region II, where Y (N) is the first operating region I and the second operating region.
The second boundary with II is shown. 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 change from the second operating region II to the first operating region II is performed.
The determination of the change of the operating range to the operating range I of the second boundary Y
It is performed based on (N). That is, the required load L is a function of the engine speed N when the engine is operating in the first operating region I and low temperature combustion is being performed.
When (N) is exceeded, it is determined that the operating region has moved to the second operating region II, and combustion is performed by the conventional combustion method. Next, when the required load L becomes lower than the second boundary Y (N) which is a function of the engine speed N, the operating region becomes the first operating region I.
It is judged that the process has been moved to No. 3, 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 load 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 load side of the second operating region II, and at this time, even if the required load L becomes lower than the first boundary X (N), low temperature combustion cannot be immediately performed. Because. That is, the low temperature combustion does not start immediately unless the required load L becomes considerably low, that is, when it becomes lower than the second boundary Y (N). The second reason is that hysteresis is provided for changes in the operating region between the first operating region I and the second operating region II.

By the way, when the engine is operating in the first operating region I and low-temperature combustion is being carried out, soot is hardly generated, and instead, unburned hydrocarbons are in a state of the above-mentioned body of soot 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 favorably oxidized by the catalyst 19 having an oxidizing function. Catalyst 1
As 9, an oxidation catalyst, a three-way reduction, 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
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 a three-way catalyst and a NO _x absorbent can be used as the catalyst 19 as described above. 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. 8 shows the throttle valve 1 for the required load L.
6, the opening of the EGR control valve 22, the EGR rate, the air-fuel ratio, the injection timing and the injection amount are shown. As shown in FIG. 8, in the first operating region I where the required load L is low, the opening degree of the throttle valve 16 is gradually increased from near full close to about half open as the required load L increases, and the EGR control valve 22 The opening degree of is gradually increased from near full close to full open as the required load L increases. Further, in the example shown in FIG. 8, the EGR rate is approximately 7 in the first operating region I.
The air-fuel ratio is set to 0% and the air-fuel ratio is slightly lean.

In other words, in the first operating region I, EGR
The opening of the throttle valve 16 and the opening of the EGR control valve 22 are controlled so that the ratio becomes approximately 70% and the air-fuel ratio becomes a lean air-fuel ratio. 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 16 is closed to near full closure, and the EGR control valve 22 is also closed to near full closure at this time. Throttle valve 1
When the valve 6 is closed to close to the fully closed state, the pressure in the combustion chamber 5 at the beginning of compression becomes low, so that 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, in idling operation, the throttle valve 16 is closed to 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 the operation range is changed from the second operation range II to the second operation range II, the opening degree of the throttle valve 16 is increased stepwise from about half open to the full open direction. At this time, in the example shown in FIG. 8, the EGR rate is reduced stepwise from approximately 70% to 40% or less, and the air-fuel ratio is increased stepwise. That is, since the EGR rate jumps over the EGR rate range (FIG. 5) in which a large amount of smoke is generated, a large amount of smoke is generated when the engine operating region changes from the first operating region I to the second operating region II. There is no.

In the second operation region II, the second combustion, that is, the combustion which is conventionally performed, is performed. Although some soot and NO _x are generated in this combustion method, the thermal efficiency is higher than that in low temperature combustion, so that 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 operation region II, the throttle valve 16 is kept fully open except for a part, and the opening degree of the EGR control valve 22 is gradually reduced as the required load L increases. Further, in this operating region II, the EGR rate becomes lower as the required load L becomes higher, and the air-fuel ratio becomes smaller as the required load L becomes higher. However, the air-fuel ratio is a lean air-fuel ratio even if the required load L becomes high.
Further, in the second operation region II, the injection start timing θS is set near the compression top dead center TDC.

FIG. 9 shows the air-fuel ratio A in the first operating region I.
/ F is shown. In FIG. 9, A / F = 15.5
Each curve shown by A / F = 16, A / F = 17, A / F = 18 shows the case where the air-fuel ratio is 15.5, 16, 17, and 18, respectively, and the air-fuel ratio between each curve is shown. Is determined by proportional distribution. As shown in FIG. 9, the first operating region I
The air-fuel ratio is lean, and the air-fuel ratio A / F is leaner as the required load L is lower in the first operating region I.

That is, the lower the required load L, the smaller the amount of heat generated by combustion. Therefore, as the required load L decreases, low temperature combustion can be performed even if the EGR rate is decreased.
If the EGR rate is reduced, the air-fuel ratio becomes large, and as shown in FIG. 9, the air-fuel ratio A / F is made larger as the required load L becomes lower. The fuel consumption rate increases as the air-fuel ratio A / F increases. Therefore, in order to make the air-fuel ratio as lean as possible, the air-fuel ratio A / F is increased as the required load L decreases in the embodiment of the present invention.

On the other hand, in the first operating region I, the EGR rate is set to the optimum EGR rate according to the operating region of the engine. This E
The GR rate EG is previously stored in the ROM 42 in the form of a map as a function of the required load L and the engine speed N as shown in FIG.
It is stored in. The target opening degree ST of the throttle valve 16 required to set the air-fuel ratio to the target air-fuel ratio A / F shown in FIG. 9 and the EGR rate to the target EGR rate EG shown in FIG. 10 is shown in FIG. ROM as a function of required load L and engine speed N as shown in FIG.
42, the air-fuel ratio is the target air-fuel ratio A / F shown in FIG. 9, and the EGR rate is the target EGR shown in FIG.
Target opening degree S of the EGR control valve 22 required to obtain the rate EG
E is stored in advance in the ROM 42 in the form of a map as a function of the required load L and the engine speed N as shown in FIG.

Further, the injection amount Q in the first operating region I
Increases as the required load L increases as shown in FIG. This injection amount Q is also a function of the engine speed,
This injection quantity Q is a function of the required load L and the engine speed N, as shown in FIG.
It is stored in 2. On the other hand, when the first combustion is performed, as described at the beginning, there is an optimum crank angle range in which stable combustion can be obtained in which smoke does not occur and misfire does not occur with respect to the fuel injection timing. When performing the first combustion, it is necessary to perform fuel injection within this optimum crank angle range. However, the optimum crank angle range in which stable combustion is obtained is the value of the operating parameter of the engine that influences the temperature of the fuel during combustion in the combustion chamber 5 and the gas temperature around it, such as the air-fuel ratio and EG.
It changes depending on the R ratio, the intake air temperature, the EGR gas temperature, the engine cooling water temperature, the humidity in the intake air, and the like.

In this case, when the value of the operating parameter of the engine changes drastically, even if the temperature of the fuel during combustion in the combustion chamber 5 and the temperature of the gas around it change, the injection timing is made to follow that change due to detection delay and the like. Cannot be changed,
As a result, the injection timing deviates from the optimum crank angle range, resulting in smoke or misfire.

By the way, among the operating parameters of the engine that affect the temperature of the fuel and the gas around it during combustion in the combustion chamber 5, the intake air temperature, the engine cooling water temperature, and the intake air humidity may suddenly change. Therefore, even if these parameters change, the injection timing can be controlled within the optimum crank angle range by following the changes. Further, as can be seen from FIG. 9, when the first combustion is being performed, the air-fuel ratio A / F does not change so much even when the operating state is changed, and similarly, when the first combustion is being performed, the operating state is changed. Changes, the EGR rate does not change so much. Therefore, even if these parameters change, the injection timing can be controlled within the optimum crank angle range by following the changes.

However, among the operating parameters of the engine that affect the temperature of the fuel and the gas around it during combustion in the combustion chamber 5, the EGR gas temperature changes drastically when the operating state of the engine changes. Therefore, even if the injection timing is controlled according to the change in the EGR gas temperature when the EGR gas temperature changes, the injection timing cannot be changed by following the change in the EGR gas temperature. Therefore, there is a problem that smoke is generated or misfire occurs due to deviation from the crank angle range.

Therefore, in the present invention, the cooling device 23 arranged in the EGR passage 21 has a cooling capacity necessary for maintaining the temperature of the EGR gas flowing out from the cooling device 23 at a substantially predetermined set temperature. I am trying to make it. In other words, supplying a large amount of engine cooling water to the cooling device 23,
Thereby, the EGR gas temperature at the outlet of the cooling device 23 is maintained at a preset temperature.

FIG. 13 shows changes in the EGR gas temperature T in the cooling device 23 shown in FIG. When the cooling device 23 shown in FIG. 1 is used as shown in FIG.
The EGR gas temperature T at the inlet of 3 is 150 ° C., 300 ° C. or 500 ° C., that is, even if the exhaust gas temperature discharged from the engine fluctuates drastically.
The EGR gas temperature T at the outlet of can be maintained at a constant temperature.

In the cooling device 13 shown in FIG. 1, the temperature of the cooling water of the engine at the outlet of the cooling device 13 is almost equal. The cooling water temperature of the engine is maintained at a constant temperature by a thermostat or the like, and in the internal combustion engine shown in FIG. 1, the cooling water temperature of the engine is maintained at approximately 120 ° C. Therefore, the EGR gas temperature T at the outlet of the cooling device 13 shown in FIG. 1 is maintained at about 120 ° C. Cooling water temperature of engine is 12
It is also possible to maintain the temperature lower than 0 ° C., for example, 80 ° C., and at this time, the EGR gas temperature T at the outlet of the cooling device 13 becomes 80 ° C.

In any case, a large amount of cooling water is supplied to the cooling device 2.
If it is supplied to 3, the cooling device 23 can be used without special temperature.
The EGR gas temperature T at the outlet of can be maintained constant. As shown in FIG. 8, the injection start timing θS1 when the first combustion is performed, that is, the injection start timing θS1 in the first operation region I is delayed as the required load L increases. That is, when the injection start timing θS1 is represented by an advance amount, the advance amount θS1 of the injection start timing becomes smaller as the required load L becomes higher as shown in FIG. 14 (A). Further, the advance amount θS1 of the injection start timing is shown in FIG.
As shown in (3), it increases as the engine speed N increases. In the present invention, when the EGR gas temperature flowing out from the cooling device 13 is a constant temperature, for example, 120 ° C., the optimum advance angle amount θS1 of the injection start timing is experimentally obtained, and the injection start timing is determined based on this advance angle amount θS1. Is determined.

Specifically, in the embodiment according to the present invention,
The EGR gas temperature at the outlet of the cooling device 13 is set in advance.
Air-fueled at a set temperature, for example 120 ° C
The ratio A / F is the standard air-fuel ratio (A / F₀), And the EGR rate E
G is the reference EGR rate EG₀And the intake air temperature GT is the standard
Temperature GT₀And the engine cooling water temperature TW is the reference temperature TW. ₀
And the humidity DF of the intake air is the reference humidity DF₀Is
The standard advance amount θS1 of the injection start timing of
The reference advance amount θS1 is shown in FIG. 14 (C).
As a function of the required load L and the engine speed N
It is stored in advance in the ROM 42 in the form of a block.

FIG. 15 shows the target air-fuel ratio when the second combustion, that is, the normal combustion by the conventional combustion method is performed. Note that in FIG. 15, A / F = 24, A / F = 3
5, each curve shown by A / F = 45 and A / F = 60 shows the target air-fuel ratios 24, 35, 45, 60, respectively. Throttle valve 1 required to set the air-fuel ratio to this target air-fuel ratio
The target opening ST of 6 is stored in advance in the ROM 42 in the form of a map as a function of the required load L and the engine speed N as shown in FIG. 16 (A), and the air-fuel ratio is set to this target air-fuel ratio. Target opening degree SE of the EGR control valve 22 required for
Is stored in advance in the ROM 42 in the form of a map as a function of the required load L and the engine speed N as shown in FIG. 16 (B).

Further, the injection quantity Q when the second combustion is performed is stored in advance in the ROM 42 in the form of a map as a function of the required load L and the engine speed N as shown in FIG. The injection start timing θS2 when the second combustion is performed is stored in advance in the ROM 42 in the form of a map as a function of the required load L and the engine speed N, as shown in FIG.

By the way, when the first combustion is performed
The air-fuel ratio A / F is the reference air-fuel ratio (A / F ₀) And EGR
The rate EG is the reference EGR rate EG₀And the intake air temperature GT is
Reference temperature GT₀And the engine cooling water temperature TW is the reference temperature T
W₀And the humidity DF of the intake air is the reference humidity DF₀And
14C, the advance amount of the injection start timing
Smoke occurs if the reference advance amount θS1
It is possible to obtain stable combustion without misfire.
Wear.

However, for example, when the intake air temperature GT into the combustion chamber 5 becomes much higher than the reference temperature GT ₀ , if the advance amount of the injection start timing is set to the reference advance amount Sθ1, The temperature of the fuel and the gas around it will be too high, resulting in smoke. Therefore, in this case, it is necessary to correct the advance amount of the injection start timing to be smaller than the reference advance amount Sθ1, that is, to delay the injection start timing.

On the other hand, when the intake air temperature GT into the combustion chamber 5 becomes much lower than the reference temperature GT ₀ , the advance amount of the injection start timing is set to the reference advance amount Sθ1. The temperature of the fuel and the gas around it during combustion will be low, resulting in misfire. Therefore, in this case, it is necessary to correct the advance amount of the injection start timing to be larger than the reference advance amount Sθ1, that is, to advance the injection start timing.

Next, the correction amount of the advance amount Sθ1 of the injection start timing will be described with reference to FIG. FIG. 19A shows the relationship between the air-fuel ratio A / F and the correction amount Δθ1 with respect to the advance amount Sθ1 of the injection start timing. When the air-fuel ratio A / F becomes large, the combustion in the combustion chamber 5 becomes active, so that the temperature of the fuel and the gas around it become high, and when the air-fuel ratio A / F becomes small, the combustion gradually becomes less active. The temperature of the fuel and the gas around it during combustion is low. Therefore, when the air-fuel ratio A / F becomes larger than the reference value (A / F) ₀ , the correction amount Δθ1 becomes a gradually larger negative value,
When the air-fuel ratio A / F becomes smaller than the reference value (A / F) ₀ , the correction amount Δθ1 gradually becomes a positive value.

FIG. 19B shows the EGR rate EG and the injection start time.
The relationship between the correction amount Δθ2 and the advance amount Sθ1
ing. If the EGR rate EG becomes large, the
Combustion gradually becomes less active.
And the gas temperature around it are low, and the EGR rate EG is small.
Combustion becomes more active when
The temperature of the gas around is high. Therefore, the EGR rate EG is
Quasi value EG₀Correction amount Δθ2 becomes larger
It becomes a positive value and the EGR rate EG becomes the reference value EG. ₀than
When it becomes smaller, the correction amount Δθ2 becomes a gradually larger negative value.
It

FIG. 19C shows the relationship between the intake air temperature GT flowing into the combustion chamber 5 and the correction amount Δθ3 with respect to the advance amount Sθ1 of the injection start timing. When the intake air temperature GT is high, the temperature of the fuel and the gas around it during combustion are high, and when the intake gas temperature GT is low, the temperature of the fuel and the gas around it are low during combustion. Therefore, when the intake air temperature GT becomes higher than the reference value GT ₀ , the correction amount Δθ3 gradually becomes a negative value, and when the intake gas temperature GT becomes lower than the reference value GT ₀ , the correction amount Δθ3 becomes gradually larger positive value.

FIG. 19D shows the relationship between the engine cooling water temperature WT and the correction amount Δθ4 with respect to the advance amount Sθ1 of the injection start timing. When the engine cooling water temperature WT rises, the gas temperature in the combustion chamber 5 rises, so the temperature of the fuel during combustion and its surrounding gas rises, and when the engine cooling water temperature WT drops, the gas temperature in the combustion chamber 5 falls. Therefore, the temperature of the fuel and the gas around it during combustion becomes low. Therefore, when the engine cooling water temperature WT becomes higher than the reference value WT ₀ , the correction amount Δθ4 gradually becomes a negative value, and the engine cooling water temperature WT becomes the reference value W.
When it becomes lower than T ₀ , the correction amount Δθ4 gradually becomes a positive value.

FIG. 19E shows the relationship between the humidity DF of the intake air and the correction amount Δθ5 with respect to the advance amount Sθ1 of the injection start timing. When the humidity DF of the intake air becomes high, the temperature of the fuel and the gas around it become low due to the endothermic action of the moisture contained in the intake air, and the humidity DF of the intake air becomes low.
The lower the temperature, the lower the endothermic action of water contained in the intake air, and the higher the temperature of the fuel and the gas around it during combustion. Therefore, the humidity DF of the intake air is the reference value DF ₀
The correction amount Δθ5 becomes a positive value that gradually increases, and the correction amount Δθ5 becomes a negative value that gradually increases when the humidity DF of the intake air becomes smaller than the reference value DF ₀ .

The relationships shown in FIGS. 19A to 19E are stored in the ROM 42 in advance. Next in FIG.
The operation control will be described with reference to FIGS. Referring to FIGS. 20 and 21, first, at step 100, the operating state of the engine is the first operating region I.
It is determined whether or not the flag I indicating that is is set. When the flag I is set, that is, when the operating state of the engine is in the first operating region I, the routine proceeds to step 101, where it is judged if the required load L has become larger than the first boundary X1 (N). To be done. L ≦ X1 (N)
In case of, the routine proceeds to step 103, where low temperature combustion is performed.

That is, at step 103, the target opening ST of the throttle valve 16 is calculated from the map shown in FIG. 11A, and the opening of the throttle valve 16 is made this target opening ST. Next, at step 104, the target opening degree SE of the EGR control valve 22 is calculated from the map shown in FIG.
The opening degree of the EGR control valve 22 is set to this target opening degree SE.
Next, at step 105, the injection amount Q is calculated from the map shown in FIG. Then, in step 106, FIG.
From the map shown in (C), the basic advance amount θS of the injection start timing
1 is calculated.

Next, at step 107, the mass flow rate Ga of the intake air detected by the mass flow rate detector 31 is fetched. Next, at step 108, the air-fuel ratio A / F is calculated from the fuel injection amount Q and the intake air mass flow rate Ga. Next, at step 109, based on this air-fuel ratio A / F, as shown in FIG.
The correction amount Δθ1 is calculated from the relationship shown in 9 (A). Next, at step 110, the correction value Δθ2 is calculated from the relationship shown in FIG. 19B based on the EGR rate EG calculated from the map shown in FIG. Then step 11
In 1, the intake air temperature GT detected by the temperature sensor 32
The correction amount Δθ3 is calculated from the relationship shown in FIG. Next, at step 112, the water temperature sensor 30
19 based on the engine cooling water temperature WT detected by FIG.
The correction value Δθ4 is calculated from the relationship shown in (D). Next, at step 113, the correction value Δθ5 is calculated from the relationship shown in FIG. 19 (E) based on the humidity of the intake air detected by the humidity sensor 33.

Next, at step 114, the basic advance angle amount θS
By adding each correction value Δθ1 to Δθ5 to 1, the final advance amount θS1 (= θS1 + Δθ of the injection start timing)
1 + Δθ2 + Δθ3 + Δθ4 + Δθ5) is calculated.
Next, at step 115, this final advance angle θS1,
The injection completion timing θE1 based on the injection amount Q and the fuel pressure in the common rail 27 detected by the fuel pressure sensor 29.
Is calculated.

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, step 118
Then, the target opening degree ST of the throttle valve 16 is calculated from the map shown in FIG. 16 (A), and the opening degree of the throttle valve 16 is made this target opening degree ST. Next, at step 119, the target opening degree SE of the EGR control valve 22 is calculated from the map shown in FIG. 16 (B), and the opening degree of the EGR control valve 22 is made this target opening degree SE. Next, at step 120, the injection amount Q is calculated from the map shown in FIG. 17, and then step 1
In 21, the injection start timing θS2 is calculated from the map shown in FIG. Next, at step 122, the injection start timing θ
The injection completion timing θE2 is calculated from S2, the injection amount Q, and the fuel pressure in the common rail 27.

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.

FIG. 22 shows another embodiment. In this embodiment, a switching device 60 for switching the flow direction of the EGR gas in the cooling device 23 to the reverse direction is provided. That is, the EGR passage portion 21a connected to the exhaust manifold 17 is bifurcated and is connected to one end A of the cooling device 23 on the one hand and to the other end B of the cooling device 23 on the other hand. A switching valve 62 which is operated by an actuator 61 is arranged at a branch portion of the EGR passage portion 21a, and the EGR passage portion 21a is selectively connected to one end A or the other end B of the cooling device 23 by the switching valve 62.

On the other hand, the EGR passage portion 21b connected to the surge tank 13 is also bifurcated and is connected to one end A of the cooling device 23 on the one hand and to the other end B of the cooling device 23 on the other hand. At the branch portion of the EGR passage portion 21b, a switching valve 6 operated by an actuator 61.
3, the EGR passage portion 21b is selectively connected to the one end A or the other end B of the cooling device 23 by the switching valve 63.

FIG. 23 shows a switching control routine for the switching valves 62 and 63. Referring to FIG. 23, step 20
At 0, the engine speed N is added to ΣN. Therefore, this Σ
N represents the cumulative value of the engine speed N. Next, at step 201, it is judged if the cumulative value ΣN of the engine speed exceeds a preset value N ₀ . When ΣN ≦ N ₀ , the processing cycle is completed.

At this time, if the switching valves 62, 63 are in the positions shown by the solid lines in FIG. 22, the switching valves 62, 63 are held at the positions shown by the solid lines in FIG. At this time, the EGR gas flows as shown by the solid arrow, and the EGR gas flows into the cooling device 23 from one end A and is discharged from the other end B. On the other hand, when it is determined at step 201 in FIG. 23 that ΣN> N ₀ , the routine proceeds to step 202, where the switching valves 62 and 63 are switched to the positions shown by the broken lines in FIG. Next, the routine proceeds to step 203, where ΣN is made zero.
At this time, the EGR gas flows as shown by the dashed arrow,
The EGR gas flows into the cooling device 23 from the other end B and is discharged from one end A.

Thereafter, in step 201 again, ΣN
When it is determined that> N ₀ , the routine proceeds to step 202, where the switching valves 62 and 63 are switched to the positions shown by the solid line in FIG. Therefore, the EGR gas is allowed to flow in the cooling device 23 in the opposite direction every time a predetermined period of time elapses. That is, when the first combustion is being performed, a large amount of unburned hydrocarbons is discharged from the engine, and thus the EGR gas contains a large amount of unburned hydrocarbons. The EGR gas is cooled by the cooling device 23, and has a considerably low temperature near the outlet of the cooling device 23. When the EGR gas temperature decreases, unburned hydrocarbons contained in the EGR gas are likely to be deposited as a deposit, and thus the deposit is deposited near the outlet of the cooling device 23.

Therefore, in this embodiment, when a predetermined period of time has passed, the EGR gas is made to flow in the cooling device 13 in the opposite direction to that in the past, and the deposit accumulated by the high temperature EGR gas is burned and removed. I am trying to do it.

[0096]

EFFECT OF THE INVENTION Smoke does not occur, and stable combustion without misfire can be obtained.

[Brief description of 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 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 opening of a throttle valve and the like.

FIG. 9 is a diagram showing an air-fuel ratio in a first operating region I.

10 is a diagram showing a map of an EGR rate in a first operating region I. FIG.

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

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

FIG. 13 is a diagram showing an EGR gas temperature in the cooling device.

FIG. 14 is a diagram showing an injection start timing.

FIG. 15 is a diagram showing an air-fuel ratio in the second combustion.

FIG. 16 is a diagram showing a map of a target opening degree of a throttle valve or the like.

FIG. 17 is a diagram showing a map of an injection amount.

FIG. 18 is a diagram showing a map of injection start timing.

FIG. 19 is a diagram showing each correction amount.

FIG. 20 is a flowchart for controlling the operation of the engine.

FIG. 21 is a flowchart for controlling the operation of the engine.

FIG. 22 is a diagram showing another embodiment of the compression ignition type internal combustion engine.

FIG. 23 is a flowchart for controlling a switching valve.

[Explanation of symbols]

6 ... Fuel injection valve 21 ... EGR passage 22 ... EGR control valve 23 ... Cooling device

─────────────────────────────────────────────────── ─── Continuation of front page (51) Int.Cl. ⁷ Identification code FI theme code (reference) F01P 7/16 504 F01P 7/16 504Z F02D 21/08 301 F02D 21/08 301D 41/14 330 41/14 330B 41/40 41/40 E 43/00 301 43/00 301J 301N 45/00 301 45/00 301F 368 368F F02M 25/07 550 F02M 25/07 550F 570 570D 580 580E F Term (reference) 3G062 AA01 AA06 BA02 BA04 BA05 BA06 CA03 CA07 CA08 EA10 ED08 FA05 FA06 GA00 GA01 GA04 GA06 GA08 GA10 GA12 GA17 3G084 AA01 AA04 BA05 BA09 BA13 BA14 BA15 BA20 CA03 CA04 DA10 DA28 EA11 EB08 EC03 FA00 FA02 FA07 FA09 FA13 FA37 FA03 FA13 FA37 FA02 FA33 FA02 FA33 FA02 FA02 FA33 FA02 FA02 FA33 FA02 FA33 FA02 FA33 FA03 FA02 FA33 FA03 FA02 CB02 CB03 CB07 DB10 EA00 EA03 EA05 EA07 EA16 EA17 EA20 FA12 FA13 FA14 FB10 FB12 GB02Y GB03Y GB04Y GB06Y GB10Y HB05 3G092 AA02 AA09 BA01 BA06 BA07 BB01 BB06 BB08 DC03 DC09 DE06S DG08 EA05 EA07 EA09 EA17 HA07 HA07 HD07Z07HA01 HAZZA01 HAZZA01 HAZZA01 HA01ZAHAZZ HA01Z HA01Z HA01Z HA01Z HA01Z HA01Z HA01Z HA01Z HA01Z HA01Z HA01Z HA01ZAHAHAZ JA23 JA24 KA07 KA08 KA09 LA00 LA03 LB13 MA01 MA11 MA18 MA20 NA08 NC01 NC02 NE12 NE13 NE15 NE23 PA00Z PA01Z PA10Z PA11A PA17Z PB08A PD01A PD11Z PD15A PD15Z PE01Z PE08Z PF03Z

Claims (15)

[Claims] 1. A soot is provided with an exhaust gas recirculation device for recirculating exhaust gas discharged from an engine into an engine intake passage, and soot increases as the amount of recirculated exhaust gas supplied to a combustion chamber increases. In an internal combustion engine in which the amount of soot generated gradually increases and reaches a peak, the amount of recirculated exhaust gas supplied to the combustion chamber is made larger than the amount of recirculated exhaust gas at which the amount of soot generated peaks. A cooling device is arranged in the exhaust gas recirculation passage that connects to the intake passage, and is required to maintain the temperature of the recirculated exhaust gas flowing out of the cooling device at a substantially preset temperature with respect to the cooling device. The optimum injection timing when the temperature of the recirculated exhaust gas flowing out from the cooling device is at the set temperature, and the injection timing is determined based on the stored injection timing. like Internal combustion engine. 2. Engine cooling water is guided to the cooling device,
The recirculated exhaust gas is cooled by the engine cooling water.
Internal combustion engine according to. 3. The internal combustion engine according to claim 2, wherein the cooling water temperature in the engine body is maintained substantially constant, and is substantially equal to the cooling water temperature in the engine body in which the set temperature is maintained substantially constant. 4. A detection means for detecting a value of an operating parameter of an engine which influences a temperature of a fuel and a gas around the fuel during combustion in a combustion chamber, the smoke is provided for the injection timing based on the value of the operating parameter. The internal combustion engine according to claim 1, wherein the control is performed between an injection timing at which the fuel injection occurs and an injection timing at which a misfire occurs. 5. The internal combustion engine according to claim 4, wherein the operating parameter is an air-fuel ratio, and the injection timing is delayed as the air-fuel ratio increases. 6. The internal combustion engine according to claim 4, wherein the operating parameter is an exhaust gas recirculation rate, and the injection timing is advanced as the exhaust gas recirculation rate increases. 7. The internal combustion engine according to claim 4, wherein the operating parameter is the intake air temperature, and the injection timing is delayed as the intake air temperature increases. 8. The internal combustion engine according to claim 4, wherein the operating parameter is a cooling water temperature of the engine, and the injection timing is delayed as the cooling water temperature of the engine increases. 9. The internal combustion engine according to claim 4, wherein the operating parameter is the humidity of the intake air, and the injection timing is advanced as the humidity of the intake air increases. 10. A switching means for switching the flow direction of the recirculated exhaust gas in the cooling device to the opposite direction,
2. The internal combustion engine according to claim 1, wherein the flow direction of the recirculated exhaust gas in the cooling device is switched to the opposite direction every time a predetermined period has elapsed. 11. The internal combustion engine according to claim 1, wherein the exhaust gas recirculation rate is approximately 55% or more. 12. The internal combustion engine according to claim 1, wherein a catalyst having an oxidizing function is arranged in the engine exhaust passage. 13. The catalyst is an oxidation catalyst, a three-way catalyst or NO.
An internal combustion engine according to claim 12, wherein the _x absorbent comprises at least one. 14. The first combustion in which the amount of recirculated exhaust gas supplied to the combustion chamber is larger than the amount of recirculated exhaust gas at which the soot generation peaks and the soot is hardly generated, and the soot generation peaks. 2. The internal combustion engine according to claim 1, further comprising switching means for selectively switching between the second combustion in which the amount of recirculated exhaust gas supplied to the combustion chamber is smaller than the amount of recirculated gas that becomes 15. 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 the first combustion is performed in the first operating region to perform the second operating region. Second in the area
15. The internal combustion engine according to claim 14, wherein the combustion is performed.