JP2000145548A - Multicylinder internal combustion engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To prevent the emission of soot and NOx simultaneously and distribute EGR gas among the combustion chambers of all cylinders with precision at the same time. SOLUTION: A larger amount of EGR gas than the amount thereof that results in the peak generation of soot is fed to each combustion chamber 5 to offer low-temperature combustion with almost no generation of soot or no emission of not only soot but NOx. Separately from an intake surge tank 12, an EGR gas surge tank 60 is provided in an EGR passage 29 upstream of the confluence of an intake branch pipe 11 and an EGR branch pipe 61 to distribute EGR gas among the combustion chambers 5 of all cylinders. EGR gas is thus distributed precisely among all the cylinder combustion chambers 5 independently of intake pulsation and others.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

The present invention relates to a multi-cylinder 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 NOx. Exhaust gas, that is, EGR gas, is recirculated through the passage into the engine intake passage. In this case, the EGR gas has a relatively high specific heat, and therefore can absorb a large amount of heat.
The combustion temperature in the combustion chamber decreases as the GR gas amount increases, that is, as the EGR rate (EGR gas amount / (EGR gas amount + intake air amount)) increases. When the combustion temperature decreases, the amount of generated NOx decreases. Therefore, the higher the EGR rate, the lower the amount of generated NOx.

As described above, it has been known that the amount of generated NOx can be reduced by increasing the EGR rate. However, when the EGR rate is increased, the soot generation amount, that is, smoke, starts to increase rapidly when the EGR rate exceeds a certain limit. In this regard, it has conventionally been considered that if the EGR rate is further increased, the smoke will increase indefinitely. Therefore, the smoke starts to increase rapidly.
The GR rate is considered to be the maximum allowable limit of the EGR rate.

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

As described above, conventionally, it has been considered that the maximum allowable limit exists for the EGR rate.
If the R rate is within the range not exceeding this maximum allowable limit, NO
The amount of x and smoke was determined to be as small as possible. However, in this way EGR
Even if the rate is set so as to minimize the generation of NOx and smoke, there is a limit to the reduction of the generation of NOx and smoke, and in fact, a considerable amount of N
At present, Ox and smoke are generated.

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

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

Accordingly, if the temperature of the fuel and the surrounding gas during combustion in the combustion chamber is suppressed to a temperature below the temperature at which the growth of hydrocarbons stops halfway, soot will not be generated, and the fuel during combustion in the combustion chamber and its surroundings will not be generated. Can be suppressed to a temperature below the temperature at which the growth of hydrocarbons stops halfway, by adjusting the amount of heat absorbed by the gas around the fuel. On the other hand, hydrocarbons whose growth has stopped halfway before reaching soot can be easily purified by post-treatment using an oxidation catalyst or the like. This is the basic idea of a new combustion system.

[0009]

However, a new combustion system as described above has not been disclosed yet. Therefore, the conventional combustion system already disclosed cannot exhibit new effects based on the new combustion system described above.

Accordingly, the present invention is to accurately distribute recirculated exhaust gas into a combustion chamber of each cylinder while simultaneously preventing soot (smoke) and NOx from being discharged from an internal combustion engine. It is an object of the present invention to provide a multi-cylinder internal combustion engine capable of performing the following.

[0011]

According to the first aspect of the present invention, there is provided an exhaust gas recirculation system having a plurality of cylinders and recirculating exhaust gas discharged from a combustion chamber into an engine intake passage. Equipped with an exhaust gas recirculation device having a circulation passage, when the amount of recirculated exhaust gas supplied to the combustion chamber is increased, the amount of soot generated gradually increases and reaches a peak, and is supplied to the combustion chamber. When the amount of recirculated exhaust gas 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 recirculated exhaust gas supplied into the combustion chamber is larger than the amount of recirculated exhaust gas at which the amount of generated soot is at a peak, and combustion in which soot is hardly generated can be performed. Distributed in the combustion chamber The recirculated exhaust gas surge tank,
There is provided a multi-cylinder internal combustion engine arranged in the exhaust gas recirculation passage upstream of a junction of the engine intake passage and the exhaust gas recirculation passage.

In the internal combustion engine according to the first aspect of the invention, the 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 amount of generated soot reaches a peak is substantially less generated. By doing so, it is possible to simultaneously prevent soot (smoke) and NOx from being discharged from the internal combustion engine. Further, a recirculation exhaust gas surge tank dedicated to recirculation exhaust gas for distributing the recirculation exhaust gas into the combustion chamber of each cylinder is provided with an exhaust gas upstream of the junction between the engine intake passage and the exhaust gas recirculation passage. It is located in the gas recirculation passage. Therefore, the recirculated exhaust gas flowing into the recirculated exhaust gas surge tank can be accurately distributed to the combustion chamber of each cylinder without being affected by pulsation of the intake air flowing in the engine intake passage.

According to the second aspect of the present invention, the intake air surge tank for distributing the intake air into the combustion chamber of each cylinder is provided upstream of the junction between the engine intake passage and the exhaust gas recirculation passage. 2. The multi-cylinder internal combustion engine according to claim 1, wherein the multi-cylinder internal combustion engine is disposed in the engine intake passage on the side of the engine.

In the internal combustion engine according to the present invention, the intake air surge tank for distributing the intake air into the combustion chamber of each cylinder is located upstream of the junction between the engine intake passage and the exhaust gas recirculation passage. It is arranged in the engine intake passage. Therefore, the intake air flowing into the intake air surge tank can be accurately distributed into the combustion chamber of each cylinder without being affected by pulsation of the recirculated exhaust gas flowing in the exhaust gas recirculation passage.

According to the third aspect of the present invention, a recirculation exhaust gas control valve for controlling the amount of recirculation exhaust gas supplied into the combustion chamber is provided adjacent to the recirculation exhaust gas surge tank. 2. The multi-cylinder internal combustion engine according to claim 1, wherein the multi-cylinder internal combustion engine is provided in the exhaust gas recirculation passage upstream of the recirculation exhaust gas surge tank.

In the internal combustion engine according to the third aspect, the recirculation exhaust gas control valve for controlling the amount of recirculation exhaust gas supplied to the combustion chamber is provided adjacent to the recirculation exhaust gas surge tank. It is arranged in the exhaust gas recirculation passage upstream of the exhaust gas surge tank. That is, the recirculation exhaust gas control valve is disposed in a portion relatively close to each cylinder in the exhaust gas recirculation passage before branching for each cylinder.
Therefore, responsiveness in controlling the amount of recirculated exhaust gas supplied to each cylinder can be improved.

According to the fourth aspect of the present invention, the plurality of intake ports branched from the engine intake passage are connected to one cylinder, and the blow-by gas for discharging the blow-by gas toward the combustion chamber. A discharge hole is provided in at least one of the plurality of branched intake ports, and a junction between the engine intake passage and the exhaust gas recirculation passage forms a remaining portion of the plurality of branched intake ports. A multi-cylinder internal combustion engine according to claim 1 disposed in an intake port is provided.

According to the fourth aspect of the present invention, the blow-by gas discharge hole is provided in at least one of the plurality of branched intake ports, and the blow-by gas discharge hole is provided between the engine intake passage and the exhaust gas recirculation passage. The junction is arranged in the remaining intake port of the plurality of branched intake ports. That is, the blow-by gas and the recirculated exhaust gas do not mix in the intake port. Therefore, it is possible to prevent deposits generated by mixing the blow-by gas and the recirculated exhaust gas from adhering in the intake port.

According to the fifth aspect of the present invention, the intake flow control valve for forming the swirl is arranged in the intake port where the junction between the engine intake passage and the exhaust gas recirculation passage is not arranged. A multi-cylinder internal combustion engine according to claim 4 is provided.

In the internal combustion engine according to the present invention, the intake flow control valve for forming the swirl is disposed in the intake port where there is no junction between the engine intake passage and the exhaust gas recirculation passage. Therefore, it is possible to prevent the deposit in the recirculated exhaust gas from adhering to the intake flow control valve.

According to the invention described in claim 6, the intake flow control valve is disposed in the intake port provided with the blow-by gas discharge hole and upstream of the blow-by gas discharge hole. The multi-cylinder internal combustion engine according to the above is provided.

In the internal combustion engine according to the present invention, the intake flow control valve is disposed in the intake port provided with the blow-by gas discharge hole and upstream of the blow-by gas discharge hole.
Therefore, the deposit in the blow-by gas can be prevented from adhering to the intake flow control valve.

According to the seventh aspect of the present invention, there is provided a multi-cylinder catalyst according to the first aspect, wherein a catalyst having an oxidizing function is disposed in the engine exhaust passage for oxidizing unburned hydrocarbons discharged from the combustion chamber. An internal combustion engine is provided.

According to the eighth aspect of the present invention, there is provided the multi-cylinder internal combustion engine according to the seventh aspect, wherein the catalyst comprises at least one of an oxidation catalyst, a three-way catalyst, and a NOx absorbent.

In the internal combustion engine according to the seventh and eighth aspects, the unburned hydrocarbon discharged from the combustion chamber is oxidized in the engine exhaust passage, so that the unburned hydrocarbon is discharged from the internal combustion engine. Can be blocked.

According to the ninth aspect of the present invention, the first combustion in which the soot is hardly generated and the recirculated exhaust gas in which the generation of soot reaches a peak are supplied to the combustion chamber more than the amount of recirculated exhaust gas. Switching means for selectively switching between the second combustion in which the amount of recirculated exhaust gas is small, and the first combustion from the first combustion or the first combustion from the second combustion. The multi-cylinder internal combustion engine according to claim 1, wherein the exhaust gas recirculation rate is changed in a stepwise manner when the internal combustion engine is switched.

In the internal combustion engine according to the ninth aspect, when the first combustion is switched to the second combustion or the second combustion is switched to the first combustion, the exhaust gas recirculation rate is changed stepwise. Accordingly, it is possible to prevent the exhaust gas recirculation rate from being set to the exhaust gas recirculation rate at which the generation amount of soot becomes a peak.

According to the tenth aspect, the exhaust gas recirculation rate during the first combustion is substantially 55% or more, and the second combustion is performed. 10. The multi-cylinder internal combustion engine according to claim 9, wherein an exhaust gas recirculation rate of the internal combustion engine is approximately 50% or less.

In the internal combustion engine according to the tenth aspect, the exhaust gas recirculation rate during the first combustion is set to approximately 55
% Or less and the exhaust gas recirculation rate when the second combustion is being performed is set to be approximately 50% or less, so that the exhaust gas recirculation rate becomes a peak at which soot generation is peaked. It can be prevented from being set to the rate.

According to the eleventh aspect of the present invention, the operating range of the engine is set to the first operating range on the low load side and the second operating range on the high load side.
10. The multi-cylinder internal combustion engine according to claim 9, wherein the first combustion region is divided into the first combustion region, and the second combustion region is divided into the second combustion region. Is provided.

[0031] In the internal combustion engine according to the eleventh aspect, when the first combustion can be performed, that is, when the temperature of the fuel and the surrounding gas during combustion in the combustion chamber can be maintained lower than the soot generation temperature. However, the first combustion is performed in the first operating region on the low load side and the second combustion is performed in the second operating region on the high load side because the heat generation amount due to the combustion is limited to the low load operation in the engine. 2 is performed. Therefore, appropriate combustion can be performed according to the operation range.

[0032]

Embodiments of the present invention will be described below with reference to the accompanying drawings.

FIG. 1 shows a four-stroke compression ignition type 4 of the present invention.
1 shows one embodiment applied to a cylinder 16-valve internal combustion engine. Referring to FIG. 1, 1 is an engine body, 2 is a cylinder block, 3 is a cylinder head, 4 is a piston, 5 is a combustion chamber, 6 is an electrically controlled fuel injection valve, 7 is an intake valve, 8 is an intake port, 9 Denotes an exhaust valve, and 10 denotes an exhaust port.
The intake port 8 is connected to an intake 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 of a compressor 16 of an exhaust turbocharger 15. Linked to The intake surge tank 12 prevents the pulsation of the intake air and accurately distributes the intake air into the combustion chamber 5 of each cylinder. The intake surge tank 12 is provided with an intake branch pipe 11.
And an EGR passage 29 which will be described later. An inlet of the compressor 16 is connected to an air cleaner 18 via an air suction pipe 17, and a throttle valve 20 driven by a step motor 19 is arranged in the air suction pipe 17. A mass flow detector 21 for detecting the mass flow of the intake air is disposed in the air suction pipe 17 upstream of the throttle valve 20.

On the other hand, the exhaust port 10 is connected to the exhaust manifold 2.
The exhaust turbine 2 of the exhaust turbocharger 15 via the
3 and an outlet of the exhaust turbine 23 is connected via an exhaust pipe 24 to a catalytic converter 26 having a built-in catalyst 25 having an oxidizing function. Exhaust manifold 22
Inside, an air-fuel ratio sensor 27 is arranged. In each exhaust manifold 22 corresponding to each cylinder, an exhaust gas temperature sensor 80 for detecting a temperature of exhaust gas from each cylinder is provided.
Is arranged.

The exhaust pipe 28 connected to the outlet of the catalytic converter 26 and the intake branch pipe 11 downstream of the intake surge tank 12
Are connected to each other via an exhaust gas recirculation (hereinafter, referred to as EGR) passage 29. Intake branch pipe 11 and EGR passage 2
In the EGR passage 29 on the upstream side of the junction with the
An EGR surge tank 60 for preventing pulsation of the GR gas and accurately distributing the EGR gas into the combustion chamber 5 of each cylinder is provided. The portion of the EGR passage downstream of the EGR surge tank 60 and branched into four by the EGR surge tank 60 is hereinafter referred to as an EGR branch pipe 61. An EGR control valve 31 driven by a step motor 30 is disposed in the EGR passage 29. The EGR control valve 31 is disposed in the EGR passage 29 adjacent to the EGR surge tank 60 and upstream of the EGR surge tank 60. An intercooler 32 for cooling EGR gas flowing through the EGR passage 29 is provided in the EGR passage 29.
Is arranged. In the embodiment shown in FIG. 1, the engine cooling water is guided into the intercooler 32, and the engine cooling water cools the EGR gas.

FIG. 2 is an enlarged detailed view of the intake surge tank 12 and the EGR surge tank 60. As shown in FIG. 2, the intake air passing through the throttle valve 20 is supplied to the intake duct 1
3 and flows into the intake surge tank 12, and is then accurately distributed to each cylinder by the intake surge tank 12,
The fuel is supplied into the combustion chamber 5 of each cylinder via each intake branch pipe 11. The intake air that has passed through the EGR control valve 31 is
Flows into the EGR surge tank 60 through the passage 29,
Next, the EGR surge tank 60 accurately distributes the fuel into each cylinder, and the EGR branch pipes 61 and the corresponding intake branch pipes 11
Is supplied into the combustion chamber 5 of each cylinder.

FIG. 3 shows the intake branch pipe 11 corresponding to one cylinder.
FIG. 3 is an enlarged detailed view of a junction between the EGR branch pipe 61 and the EGR branch pipe 61. FIG.
As shown in the figure, the intake branch pipe 11 corresponding to one cylinder is
It is connected to the cylinder via intake ports 71 and 72 branched into two. A blow-by discharge hole 81 for recirculating blow-by gas, fuel vapor (evaporation gas), and the like into the intake port 72 and discharging the blow-by gas and fuel vapor into the combustion chamber 5 has one of the two branched intake ports 71, 72.
Is placed within. Here, the blow-by gas is discharged into the crankcase from the gap between the piston rings during the compression stroke and the explosion stroke of the engine, and
Through the gap between the inner wall and the outer wall.

An EGR branch pipe 61 extending from the EGR surge tank 60 is connected to two branched intake ports 7.
It merges with the other one of the intake ports 71, 72. An intake flow control valve 80 for forming a swirl is disposed in an intake port that does not merge with the EGR branch pipe 61, that is, an intake port 72 provided with a blow-by gas discharge hole 81, and is located upstream of the blow-by gas discharge hole 81. Placed on the side. Although not shown, the blow-by gas discharge holes 81
Is also connected to the purge line of the evaporation gas control system.

Returning to the description of FIG. 1, the fuel injection valve 6 is connected to a fuel reservoir, a so-called common rail 34, via a fuel supply pipe 33. Fuel is supplied into the common rail 34 from an electric control type variable discharge fuel pump 35, and the fuel supplied into the common rail 34 is supplied to the fuel injection valve 6 through each fuel supply pipe 33. A fuel pressure sensor 36 for detecting the fuel pressure in the common rail 34 is attached to the common rail 34, and the fuel pump 35 is controlled so that the fuel pressure in the common rail 34 becomes the target fuel pressure based on the output signal of the fuel pressure sensor 36. Is controlled.

The electronic control unit 40 is composed of a digital computer, and is connected to a read only memory (ROM) 42, a random access memory (RAM) 43, a CPU (microprocessor) 44, an input port 45, An output port 46 is provided. The output signal of the mass flow detector 21 is input to the input port 45 via the corresponding AD converter 47, and the output signals of the air-fuel ratio sensor 27 and the fuel pressure sensor 36 are also input to the input port via the corresponding AD converter 47, respectively. 45 is input. The output signal of the exhaust gas temperature sensor 80 disposed in each exhaust manifold 22 corresponding to each cylinder 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 amount of depression 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. . The input port 45 is connected to a crank angle sensor 52 that generates an output pulse every time the crankshaft rotates, for example, by 30 °. The engine speed is calculated based on the output value of the crank angle sensor 52. 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 a corresponding drive circuit 48.

Each exhaust gas temperature sensor 8 corresponding to each cylinder
From the output value of 0, the average value of the output values of all the exhaust gas temperature sensors 80 is calculated, and the cylinder in which the difference between the output value of the exhaust gas temperature sensor 80 and the calculated average value is equal to or more than a predetermined value, It is determined that the fuel injection amount varies as compared with the other cylinders. Next, in the cylinder in which the fuel injection amount varies, the correction for increasing or decreasing the fuel injection period is performed, and the variation in the fuel injection amount is reduced.

FIG. 4 shows the throttle valve 2 during low engine load operation.
4 shows changes in output torque and changes in smoke, HC, CO, and NOx emissions when the air-fuel ratio A / F (horizontal axis in FIG. 4) is changed by changing the opening degree and the EGR rate of zero. 7 shows an experimental example. As can be seen from FIG. 4, in this experimental example, the EGR rate increases as the air-fuel ratio A / F decreases, and when the air-fuel ratio is equal to or lower than the stoichiometric air-fuel ratio (≒ 14.6), the EGR rate is 65% or higher.

As shown in FIG. 4, when the air-fuel ratio A / F is reduced by increasing the EGR rate, the smoke is reduced when the EGR rate becomes close to 40% and the air-fuel ratio A / F becomes about 30. The generation starts to increase. Next, when the EGR rate is further increased and the air-fuel ratio A / F is reduced, the amount of smoke generated sharply increases and reaches a peak. Next, when the EGR rate is further increased and the air-fuel ratio A / F is reduced, the smoke is sharply reduced. When the EGR rate is increased to 65% or more and the air-fuel ratio A / F is around 15.0, the smoke becomes almost zero. . That is, almost no soot is generated. At this time, the output torque of the engine slightly decreases, and N
The generation amount of Ox is considerably reduced. On the other hand, at this time, HC,
The amount of generated CO starts to increase.

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

The following can be said from the experimental results shown in FIGS. That is, first, the air-fuel ratio A / F is 1
When the amount of smoke generation is almost zero at 5.0 or less, FIG.
As shown in (2), the generation amount of NOx is considerably reduced. N
The decrease in the amount of generated Ox means that the combustion temperature in the combustion chamber 5 has decreased. Therefore, it can be said that the combustion temperature in the combustion chamber 5 has decreased when little soot is generated. . The same can be said from FIG. That is, in the state shown in FIG. 5B where almost no soot is generated, the combustion pressure is low.
The combustion temperature inside is low.

Second, when the amount of generated smoke, that is, the amount of generated soot becomes almost zero, as shown in FIG.
Emissions increase. This means that hydrocarbons are emitted without growing to soot. That is, the linear hydrocarbons and aromatic hydrocarbons contained in the fuel as shown in FIG. 6 are thermally decomposed when the temperature is increased in a state of lack of oxygen to form a soot precursor, and then mainly, Soot consisting of a solid aggregate of carbon atoms is produced. In this case, the actual soot generation process is complicated, and it is not clear what form the soot precursor takes, but in any case, the hydrocarbon as shown in FIG. It will grow to soot. Therefore, as described above, when the generation amount of soot becomes almost zero, the emission amounts of HC and CO increase as shown in FIG. 4, but HC at this time is a soot precursor or a hydrocarbon in a state before it. .

When these considerations based on the experimental results shown in FIGS. 4 and 5 are summarized, when the combustion temperature in the combustion chamber 5 is low, the amount of soot generation becomes almost zero. Is discharged from the combustion chamber 5. As a result of further detailed experimental study on this, if the temperature of the fuel and the surrounding gas in the combustion chamber 5 is lower than a certain temperature, the growth process of the soot is stopped halfway, that is, the soot is It was found that no soot was generated, and soot was generated when the temperature of the fuel and its surroundings in the combustion chamber 5 exceeded a certain temperature.

The temperature of the fuel and its surroundings when the process of producing hydrocarbons is stopped in the state of the soot precursor, that is, the above-mentioned certain temperature depends on various factors such as the type of fuel and the compression ratio of the air-fuel ratio. Although it cannot be said how many times the temperature changes, this certain temperature has a deep relationship with the amount of generated NOx. Therefore, this certain temperature can be defined to some extent from the amount of generated NOx. it can. That is, as the EGR rate increases, the temperature of the fuel during combustion and the gas temperature around it decrease, and the amount of generated NOx decreases. At this time, when the generation amount of NOx becomes about 10 p.pm or less, soot is hardly generated. Therefore, the above certain temperature is NO
The temperature almost coincides with the temperature when the amount of generated x is about 10 p.pm or less.

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

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

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

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

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

FIG. 7 shows the relationship between the EGR rate and smoke when the EGR gas is used as the inert gas and the degree of cooling of the EGR gas is changed. That is, in FIG. 7, the curve A indicates that the EGR gas temperature is substantially 9
Curve B shows the case where the EGR gas is cooled by a small cooling device, and curve C shows the case where the temperature is maintained at 0 ° C.
Indicates a case where the EGR gas is not forcibly cooled.

As shown by the curve A in FIG. 7, when the EGR gas is strongly cooled, the amount of soot generation peaks at a point where the EGR rate is slightly lower than 50%. Above a percentage, little soot is generated.

On the other hand, as shown by the curve B in FIG.
When the R gas is cooled slightly, the amount of soot generation peaks when the EGR rate is slightly higher than 50%,
In this case, if the EGR rate is set to about 65% or more, almost no soot is generated.

As shown by the curve C in FIG.
When the R gas is not forcibly cooled, the EGR rate becomes 5
The soot generation amount peaks near 5%, and in this case, if the EGR rate is set to approximately 70% or more, soot is hardly generated.

FIG. 7 shows the amount of smoke generated when the engine load is relatively high. When the engine load decreases, the EGR rate at which the amount of soot peaks slightly decreases, and almost no soot is generated. The lower limit of the EGR rate to be eliminated also slightly decreases. As described above, the lower limit of the EGR rate at which almost no soot is generated varies depending on the degree of cooling of the EGR gas and the engine load.

FIG. 8 shows a mixture of EGR gas and air necessary to reduce the temperature of fuel during combustion and the surrounding gas to a temperature lower than the temperature at which soot is generated when EGR gas is used as the inert gas. It shows the gas amount, the ratio of air in the mixed gas amount, and the ratio of EGR gas in the mixed gas. In FIG. 8, the vertical axis indicates the total intake gas amount drawn into the combustion chamber 5, and the dashed line Y indicates the total intake gas amount that can be drawn into the combustion chamber 5 when supercharging is not performed. ing. The horizontal axis indicates the required load.

Referring to FIG. 8, the proportion of air, that is, the amount of air in the mixed gas, indicates the amount of air necessary to completely burn the injected fuel. That is, in the case shown in FIG. 8, the ratio between the air amount and the injected fuel amount is the stoichiometric air-fuel ratio. On the other hand, in FIG. 8, the ratio of the EGR gas, that is, the amount of the EGR gas in the mixed gas is set so that when the injected fuel is burned, the temperature of the fuel and the surrounding gas is lower than the temperature at which soot is formed. The required minimum EGR gas amount is shown. This EGR gas amount is approximately 55% or more in terms of the EGR rate, and is 70% or more in the embodiment shown in FIG. That is, the total intake gas amount sucked into the combustion chamber 5 is represented by a solid line X in FIG. 8, and the ratio between the air amount and the EGR gas amount in the total intake gas amount X is as shown in FIG. The temperature of the fuel and the gas around it will be lower than the temperature at which soot is produced, so that no soot is generated. At this time, NOx
The amount generated is around 10 p.pm or less, and therefore the amount of NOx generated is extremely small.

When the fuel injection amount increases, the calorific value when the fuel burns increases. Therefore, in order to maintain the temperature of the fuel and the surrounding gas at a temperature lower than the temperature at which the soot is generated, the heat generated by the EGR gas is required. Must be increased. Therefore, as shown in FIG. 8, the EGR gas amount must be increased as the injected fuel amount increases.
That is, the EGR gas amount needs to increase as the required load increases.

When the supercharging is not performed, the upper limit of the total intake gas amount X sucked into the combustion chamber 5 is Y. Therefore, in FIG. 8, in the region where the required load is larger than Lo, the required load is reduced. As the ratio increases, the air-fuel ratio cannot be maintained at the stoichiometric air-fuel ratio unless the EGR gas ratio is reduced. In other words, when the supercharging is not performed and the required air-fuel ratio is maintained at the stoichiometric air-fuel ratio in an area where the required load is larger than Lo, the EGR rate decreases as the required load increases, and In the region where the required load is larger than Lo, the temperature of the fuel and the surrounding gas cannot be maintained at a temperature lower than the temperature at which soot is generated.

However, as shown in FIG. 1, when the EGR gas is recirculated through the EGR passage 29 to the inlet side of the supercharger, that is, to the air suction pipe 17 of the exhaust turbocharger 15, the required load is larger than Lo. In EGR rate 5
It can be maintained at 5% or more, for example 70%, so that the temperature of the fuel and its surrounding gas can be kept below the temperature at which soot is produced. That is, if the EGR gas is recirculated so that the EGR rate in the air suction pipe 17 becomes, for example, 70%, the EGR rate of the suction gas boosted by the compressor 16 of the exhaust turbocharger 15 also becomes 70%. The temperature of the fuel and the surrounding gas can be maintained at a temperature lower than the temperature at which soot is generated, to the extent that the pressure can be increased by the compressor 16. Therefore, the operating range of the engine that can generate low-temperature combustion can be expanded. When the EGR rate is set to 55% or more in a region where the required load is larger than Lo, E
The throttle valve 20 is opened when the GR control valve 31 is fully opened.
Is slightly closed.

As described above, FIG. 8 shows the case where the fuel is burned under the stoichiometric air-fuel ratio. However, even if the air amount is smaller than the air amount shown in FIG. 8, that is, the air-fuel ratio is made rich. Even so, while suppressing the generation of soot, the generation amount of NOx was reduced to 10 p.p.
m or less, and even if the air amount is larger than the air amount shown in FIG. 8, that is, even if the average value of the air-fuel ratio is lean from 17 to 18, the generation of soot is prevented. Meanwhile, the amount of generated NOx can be reduced to about 10 p.pm or less.

That is, when the air-fuel ratio is made rich, the fuel becomes excessive, but since the combustion temperature is suppressed to a low temperature, the excess fuel does not grow to soot, thus producing soot. There is no. At this time, only a very small amount of NOx 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 increases, but in the present invention, the soot is suppressed to a low temperature, so that the soot is reduced. Not generated at all. Furthermore, NOx
Only very small amounts are generated.

As described above, when low-temperature combustion is performed, soot is generated regardless of the air-fuel ratio, that is, whether the air-fuel ratio is rich, the stoichiometric air-fuel ratio, or the average air-fuel ratio is lean. However, the generation amount of NOx becomes extremely small. Therefore, considering the improvement of the fuel consumption rate, it can be said that it is preferable to make the average air-fuel ratio lean at this time.

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

FIG. 9 shows a first operation region I in which the first combustion, that is, low-temperature combustion is performed, and a second operation region II in which the second combustion, that is, combustion by the conventional combustion method, is performed. I have. In FIG. 9, the vertical axis L indicates the amount of depression of the accelerator pedal 50, that is, the required load, and the horizontal axis N indicates the engine speed. In FIG. 9, X (N) is the first
Shows the first boundary between the operating region I and the second operating region II, and Y (N) represents the first operating region I and the second operating region.
2 shows a second boundary with II. The determination of the change of the operating range from the first operating range I to the second operating range II is made based on the first boundary X (N), and the determination of the change from the second operating range II to the first
The determination of the change of the operation region to the operation region I of the second boundary Y
(N).

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

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

By the way, when the operating region of the engine is in the first operating region I and low-temperature combustion is being performed, soot is hardly generated, but the unburned hydrocarbon is replaced with the precursor of soot or the state before it. It is discharged from the combustion chamber 5 in the form. At this time, the unburned hydrocarbon discharged from the combustion chamber 5 is oxidized well by the catalyst 25 having an oxidizing function.

As the catalyst 25, an oxidation catalyst, a three-way catalyst, or a NOx absorbent can be used. When the average air-fuel ratio in the combustion chamber 5 is lean, the NOx absorbent
And has the function of releasing NOx when the average air-fuel ratio in the combustion chamber 5 becomes rich.

This NOx absorbent uses, for example, alumina as a carrier and, for example, potassium K, sodium N
a, lithium Li, at least one selected from alkali metals such as cesium Cs, alkaline earths such as barium Ba and calcium Ca, rare earths such as lanthanum La and yttrium Y, and noble metals such as platinum Pt. Is carried.

The oxidation catalyst, the three-way catalyst and the NO
The x absorbent also has an oxidizing function, and thus the three-way catalyst and the NOx absorbent can be used as the catalyst 25 as described above.

FIG. 10 shows the output of the air-fuel ratio sensor 27. As shown in FIG. 10, 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, referring to FIG. 11, the first operation region I will be described.
The operation control in the second operation region II will be schematically described.

FIG. 11 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 load L. As shown in FIG. 11, in the first operation region I where the required load L is low, the opening of the throttle valve 20 is gradually increased from almost fully closed to about ／ 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.
The R rate is approximately 70%, and the air-fuel ratio is a slightly lean air-fuel ratio.

In other words, in the first operating region I, the EGR
The opening of the throttle valve 20 and the opening of the EGR control valve 31 are controlled such that the rate becomes approximately 70% and the air-fuel ratio becomes a slightly lean air-fuel ratio. 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 is delayed as the required load L is increased, and the injection completion timing θE is delayed as the injection start timing θS is delayed.

During idle operation, the throttle valve 2
0 is closed to almost fully closed, and at this time, the EGR control valve 31
Is also closed to near full closure. When the throttle valve 20 is closed close to the fully closed state, the pressure in the combustion chamber 5 at the start of compression decreases, so that the compression pressure decreases. When the compression pressure decreases, the compression work by the piston 4 decreases, so that the vibration of the engine body 1 decreases. That is, during idling operation, the throttle valve 2
0 is closed until it is almost fully closed.

On the other hand, the operating range of the engine is the first operating range I
From the second operating region II to the second operating region II, the opening of the throttle valve 20 is increased stepwise from about 2/3 opening toward the full opening direction. At this time, in the example shown in FIG. 11, 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 rate at which the EGR rate generates a large amount of smoke
The engine operating range is the first because it jumps over the rate range (Fig. 5).
A large amount of smoke does not occur when changing from the operating region I to the second operating region II.

In the second operating region II, the conventional combustion is performed. In the second operating region II, the throttle valve 20 is held in a fully open state except for a part, and the opening of the EGR control valve 31 is gradually reduced as the required load L increases. In this operating region II, the EGR rate decreases as the required load L increases, and the air-fuel ratio decreases as the required load L increases. However, the air-fuel ratio is a lean air-fuel ratio even when the required load L increases. In the second operation region II, the injection start timing θS is set near the compression top dead center TDC.

FIG. 12A shows the target air-fuel ratio A / F in the first operation region I. In FIG. 12A, A / F = 15.5, A / F = 16, A / F = 17,
Each curve represented by A / F = 18 has a target air-fuel ratio of 1
5.5, 16, 17, and 18, and the air-fuel ratio between the curves is determined by proportional distribution. FIG.
As shown in (A), the air-fuel ratio is lean in the first operating region I, and in the first operating region I, the target air-fuel ratio A / F becomes leaner as the required load L decreases.

That is, the lower the required load L, the smaller the amount of heat generated by combustion. Therefore, low-temperature combustion can be performed even if the EGR rate is reduced as the required load L decreases.
When the EGR rate is reduced, the air-fuel ratio increases. Therefore, as shown in FIG. 12A, as the required load L decreases, the target air-fuel ratio A / F increases. Target air-fuel ratio A /
As F increases, the fuel consumption rate increases. Accordingly, in order to make the air-fuel ratio as lean as possible, in the embodiment according to the present invention, the target air-fuel ratio A / F is increased as the required load L decreases.

The target air-fuel ratio A / F shown in FIG. 12A is stored in advance in a ROM 4 as a function of the required load L and the engine speed N as shown in FIG.
2 is stored. In addition, the throttle valve 2 required to set the air-fuel ratio to the target air-fuel ratio A / F shown in FIG.
The target opening ST of 0 is stored in the ROM 42 in advance in the form of a map as a function of the required load L and the engine speed N as shown in FIG.
EGR required to achieve target air-fuel ratio A / F shown in (A)
The target opening SE of the control valve 31 is stored in the ROM 42 in advance in the form of a map as a function of the required load L and the engine speed N as shown in FIG.

FIG. 14A shows the target air-fuel ratio A when the second combustion, that is, the normal combustion by the conventional combustion method is performed.
/ F. Note that A / F in FIG.
= 24, A / F = 35, A / F = 45, and A / F = 60 indicate target air-fuel ratios of 24, 35, 45, and 6, respectively.
0 is shown. The target air-fuel ratio A shown in FIG.
/ F is a function of the required load L and the engine speed N as shown in FIG.
Is stored within. Further, the throttle valve 20 required to set the air-fuel ratio to the target air-fuel ratio A / F shown in FIG.
As shown in FIG. 15 (A), the target opening ST is stored in the ROM 42 in advance in the form of a map as a function of the required load L and the engine speed N, and the air-fuel ratio is shown in FIG. 14 (A).
As shown in FIG. 15B, the target opening degree SE of the EGR control valve 31 required to obtain the target air-fuel ratio A / F shown in FIG. It is stored in the ROM 42.

When the second combustion is being performed, the fuel injection amount Q is calculated based on the required load L and the engine speed N. The fuel injection amount Q is stored in the ROM 42 in advance as a function of the required load L and the engine speed N as shown in FIG.

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 a flag I indicating that the operating state of the engine is in the first operating region I is set. When the flag I is set, that is, when the operating state of the engine is in the first operating region I, step 10
The process proceeds to 1 to determine 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, in step 103, the target opening ST of the throttle valve 20 is calculated from the map shown in FIG. 13A, and the opening of the throttle valve 20 is set to this target opening ST. Next, at step 104, the target opening degree SE of the EGR control valve 31 is calculated from the map shown in FIG.
The opening of the EGR control valve 31 is set as the target opening SE.
Next, at step 105, the mass flow rate of the intake air (hereinafter simply referred to as the intake air amount) Ga detected by the mass flow rate detector 21 is taken in. Next, at step 106, FIG.
The target air-fuel ratio A / F is calculated from the map shown in FIG. Next, at step 107, based on the intake air amount Ga and the target air-fuel ratio A / F, a fuel injection amount Q necessary for setting the air-fuel ratio to the target air-fuel ratio A / F is calculated.

When the required load L or the engine speed N changes during the low-temperature combustion, the opening of the throttle valve 20 and the opening of the EGR control valve 31 immediately change to the required load L and the engine speed N. Target opening ST according to
Matched to SE. Therefore, for example, when the required load L is increased, the amount of air in the combustion chamber 5 is immediately increased, and the generated torque of the engine is immediately increased.

On the other hand, the opening degree of the throttle valve 20 or the EGR
When the opening degree of the control valve 31 changes and the intake air amount changes, the change in the intake air amount Ga is detected by the mass flow rate detector 21, and the fuel injection amount Q is controlled based on the detected intake air amount Ga. Is done. That is, the fuel injection amount Q is changed after the intake air amount Ga actually changes.

When it is determined in step 101 that L> X (N), the routine proceeds to step 102, where the flag I is reset. Then, the routine proceeds to step 110, where the second combustion is performed.

That is, in step 110, the target fuel injection amount Q is calculated from the map shown in FIG. 16, and the fuel injection amount is set as the target fuel injection amount Q. Then step 11
In FIG. 1, the throttle valve 20 is obtained from the map shown in FIG.
Is calculated. Next, at step 112, the target opening SE of the EGR control valve 31 is calculated from the map shown in FIG. 15B, and the opening of the EGR control valve 31 is set as the target opening SE.

Next, at step 113, the intake air amount Ga detected by the mass flow detector 21 is taken. Next, at step 114, the fuel injection amount Q and the intake air amount Ga
From this, the actual air-fuel ratio (A / F) _R is calculated. Next, at step 115, the target air-fuel ratio A / F is calculated from the map shown in FIG. Next, at step 116, it is determined whether or not the actual air-fuel ratio (A / F) _R is larger than the target air-fuel ratio A / F. (A / F) When _R > A / F, the routine proceeds to step 117, where the throttle opening correction value ΔS
T is reduced by a constant value α, then step 11
Go to 9. On the other hand, when (A / F) _R ≤A / F, the routine proceeds to step 118, where the correction value ΔST is increased by a constant value α, and then the routine proceeds to step 119. In step 119, the final target opening ST is calculated by adding the correction value ΔST to the target opening ST of the throttle valve 20, and the opening of the throttle valve 20 is set as the final target opening ST. That is, the opening of the throttle valve 20 is controlled so that the actual air-fuel ratio (A / F) _R becomes the target air-fuel ratio A / F.

When the required load L or the engine speed N changes during the second combustion, the fuel injection amount immediately matches the target fuel injection amount Q corresponding to the required load L and the engine speed N. I'm sullen. For example, the required load L
Is increased, the fuel injection amount is immediately increased, and thus the generated torque of the engine is immediately increased.

On the other hand, when the fuel injection amount Q is increased and the air-fuel ratio deviates from the target air-fuel ratio A / F, the opening of the throttle valve 20 is controlled so that the air-fuel ratio becomes the target air-fuel ratio A / F. That is, the air-fuel ratio is changed after the fuel injection amount Q changes.

When the flag I is reset, in the next processing cycle, the process proceeds from step 100 to step 108, where it is determined whether or not the required load L has become lower than the second boundary Y (N). Step 110 when L ≧ Y (N)
And the second combustion is performed under the lean air-fuel ratio.

On the other hand, at step 108, L <Y
When it is determined that (N) has been reached, the routine proceeds to step 109, where the flag I is set, and then proceeds to step 103 to perform low-temperature combustion.

In the embodiments described above, the fuel injection amount Q is controlled by open loop when low-temperature combustion is performed, and the air-fuel ratio changes the opening of the throttle valve 20 when the second combustion is performed. It is controlled by changing.
However, the fuel injection amount Q can be feedback-controlled based on the output signal of the air-fuel ratio sensor 27 when the low-temperature combustion is being performed, and the air-fuel ratio can be controlled by the EGR control when the second combustion is being performed. The control can also be performed by changing the opening of the valve 31.

As described above, according to the present embodiment, FIG.
In step 103 to step 107 of FIG. 7, by performing low-temperature combustion in which the amount of EGR gas supplied into the combustion chamber 5 is larger than the amount of EGR gas at which the generation amount of soot is at a peak and soot is hardly generated, It is possible to simultaneously prevent the emission of soot and the emission of NOx. Further, an EGR gas dedicated to the EGR gas for distributing the EGR gas into the combustion chamber 5 of each cylinder.
The surge tank 60 is disposed in the EGR passage 29 on the upstream side of the junction of the EGR branch pipe 61 and the intake branch pipe 11. Therefore, the EGR gas flowing into the EGR surge tank 60 can be accurately distributed into the combustion chamber 5 of each cylinder without being affected by the pulsation of the intake air flowing in the engine intake passage.

Further, according to the present embodiment, the intake surge tank 12 for distributing the intake air into the combustion chamber 5 of each cylinder.
Are arranged in the engine intake passage on the upstream side of the junction of the intake branch pipe 11 and the EGR branch pipe 61. Therefore, EGR
The intake air flowing into the intake surge tank 12 can be accurately distributed into the combustion chamber 5 of each cylinder without being affected by pulsation of the EGR gas flowing in the passage 29.

Further, according to the present embodiment, the EGR control valve 31 for controlling the amount of EGR gas supplied into the combustion chamber 5 is provided adjacent to the EGR surge tank 60 and upstream of the EGR surge tank 60. In the EGR passage 29. That is, the EGR control valve 31 is arranged in a portion relatively close to each cylinder in the EGR passage 29 before branching for each cylinder. Therefore, EG supplied to each cylinder
Responsiveness when controlling the R gas amount can be improved.

Further, according to this embodiment, the blow-by gas discharge hole 81 is provided in one of the branched intake ports 71, 72, and the intake branch pipe 11 and the EGR branch pipe 61 are provided. Is disposed in the other intake port 71. That is, the blow-by gas and the EGR gas do not mix in the intake port 71 or 72. Therefore, a deposit generated by mixing the blow-by gas and the EGR gas is generated in the intake port 71.
Or, it can be avoided that it adheres in the inside 72.

Further, according to the present embodiment, the intake flow control valve 80 for forming swirl is disposed in the intake port 72 where there is no junction between the intake branch pipe and the EGR branch pipe. Therefore, it is possible to prevent the deposit in the EGR gas from adhering to the intake flow control valve 80.

Further, according to the present embodiment, the intake flow control valve 80 is arranged in the intake port 72 provided with the blow-by gas discharge hole 81 and upstream of the blow-by gas discharge hole 81. Therefore, it is possible to prevent the deposit in the blow-by gas from adhering to the intake flow control valve 80.

[0105]

According to the first aspect of the present invention, the recirculated exhaust gas is combusted in each cylinder while simultaneously preventing soot (smoke) and NOx from being discharged from the internal combustion engine. It can be accurately distributed in the room.

According to the second aspect of the present invention, the intake air flowing into the intake air surge tank is subjected to the combustion of each cylinder without being affected by the pulsation of the recirculated exhaust gas flowing in the exhaust gas recirculation passage. It can be accurately distributed in the room.

According to the third aspect of the present invention, the responsiveness in controlling the amount of recirculated exhaust gas supplied to each cylinder can be improved.

According to the fourth aspect of the present invention, it is possible to prevent deposits generated by mixing blow-by gas and recirculated exhaust gas from adhering to the intake port.

According to the fifth aspect of the present invention, it is possible to prevent deposits in the recirculated exhaust gas from adhering to the intake flow control valve.

According to the sixth aspect of the present invention, it is possible to prevent the deposit in the blow-by gas from adhering to the intake flow control valve.

According to the seventh and eighth aspects of the present invention, it is possible to prevent unburned hydrocarbons from being discharged from the internal combustion engine.

According to the ninth and tenth aspects of the present invention,
The exhaust gas recirculation rate can be prevented from being set to the exhaust gas recirculation rate at which the generation amount of soot becomes a peak.

According to the eleventh aspect, appropriate combustion can be performed according to the operation range.

[Brief description of the drawings]

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

FIG. 2 is an enlarged detailed view of an intake surge tank and an EGR surge tank.

FIG. 3 is an enlarged detailed view of a junction of an intake branch pipe and an EGR branch pipe corresponding to one cylinder.

FIG. 4 is a diagram showing amounts of smoke and NOx generated;

FIG. 5 is a diagram showing a combustion pressure.

FIG. 6 is a diagram showing fuel molecules.

FIG. 7 is a diagram showing a relationship between a generation amount of smoke and an EGR rate.

FIG. 8 is a diagram showing a relationship between an injection fuel amount and a mixed gas amount.

FIG. 9 is a diagram showing a first operation region I and a second operation region II.

FIG. 10 is a diagram showing an output of an air-fuel ratio sensor.

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

FIG. 12 is a diagram showing an air-fuel ratio and the like in a first operation region I.

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

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

FIG. 15 is a view showing a map of a target opening degree of a throttle valve and the like.

FIG. 16 is a diagram showing a map of a fuel injection amount.

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

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 5 ... Combustion chamber 6 ... Fuel injection valve 11 ... Intake branch pipe 12 ... Intake surge tank 29 ... EGR passage 31 ... EGR control valve 60 ... EGR surge tank 61 ... EGR branch pipe

Continued on the front page (51) Int.Cl. ⁷ Identification symbol FI Theme coat II (reference) F01N 3/08 F01N 3/08 A 3/24 3/24 SE F02B 31/02 F02B 31/02 J F02D 21/08 F02D 21/08 L F02M 35/10 311 F02M 35/10 311E (72) Inventor Masato Goto 1 Toyota Town, Toyota City, Aichi Prefecture F-term in Toyota Motor Corporation (reference) 3G015 AA07 AA13 BD16 FA02 3G062 AA01 AA03 AA05 BA01 BA02 DA01 EA11 ED01 ED04 ED08 GA04 GA06 3G091 AA11 AA18 AA28 AB01 BA14 BA19 EA01 EA03 EA07 FA07 FA13 FA14 GB02W GB03W GB04W GB06W HA03 HA36 HB05 HB06 HB09 3G092 AA02 AA19 AA18 AA18 AA18 AA18 AA18 AA18 AA13 HF08Z

Claims (11)

[Claims] An exhaust gas recirculation device including a plurality of cylinders and an exhaust gas recirculation passage for recirculating exhaust gas discharged from a combustion chamber into an engine intake passage; As the amount of recirculated exhaust gas supplied into the chamber increases, the amount of soot generated gradually increases and reaches a peak, and as the amount of recirculated exhaust gas supplied into the combustion chamber further increases, combustion occurs The amount of recirculated exhaust gas in a multi-cylinder internal combustion engine in which the temperature of fuel during combustion in a room and its surrounding gas is lower than the soot generation temperature and soot is hardly generated, and the amount of soot generation peaks It is possible to perform combustion in which the amount of recirculated exhaust gas supplied to the combustion chamber is large and little soot is generated, and a recirculated exhaust gas surge tank for distributing the recirculated exhaust gas to the combustion chamber of each cylinder is provided. , Engine intake passage A multi-cylinder internal combustion engine disposed in the exhaust gas recirculation passage upstream of the junction of the exhaust gas recirculation passage and the exhaust gas recirculation passage. 2. An intake air surge tank for distributing intake air into a combustion chamber of each cylinder is disposed in the engine intake passage upstream of a junction between the engine intake passage and the exhaust gas recirculation passage. The multi-cylinder internal combustion engine according to claim 1. 3. A recirculation exhaust gas control valve for controlling an amount of recirculation exhaust gas supplied to a combustion chamber is provided adjacent to the recirculation exhaust gas surge tank and higher than the recirculation exhaust gas surge tank. 2. The multi-cylinder internal combustion engine according to claim 1, wherein the internal combustion engine is disposed in the exhaust gas recirculation passage on the upstream side. 4. A plurality of intake ports branched from an engine intake passage are connected to one cylinder, and a plurality of blow-by gas discharge holes for discharging blow-by gas into a combustion chamber are provided. The at least one intake port, wherein a junction of the engine intake passage and the exhaust gas recirculation passage is disposed in the remaining intake port of the plurality of branched intake ports. 2. The multi-cylinder internal combustion engine according to 1. 5. The multi-cylinder engine according to claim 4, wherein an intake flow control valve for forming a swirl is arranged in an intake port where a junction between the engine intake passage and the exhaust gas recirculation passage is not arranged. Internal combustion engine. 6. The multi-cylinder internal combustion engine according to claim 5, wherein the intake flow control valve is disposed in an intake port provided with the blow-by gas discharge hole and upstream of the blow-by gas discharge hole. 7. The multi-cylinder internal combustion engine according to claim 1, wherein a catalyst having an oxidation function is disposed in an engine exhaust passage for oxidizing unburned hydrocarbons discharged from the combustion chamber. 8. The catalyst according to claim 1, wherein the catalyst is an oxidation catalyst, a three-way catalyst or NO.
A multi-cylinder internal combustion engine according to claim 7, comprising at least one of the x absorbents. 9. The first combustion in which the soot is hardly generated and the amount of recirculated exhaust gas supplied into the combustion chamber is smaller than the amount of recirculated exhaust gas in which the amount of generated soot is peaked. Switching means for selectively switching between a small amount of second combustion and exhaust gas recirculation when switching from the first combustion to the second combustion or from the second combustion to the first combustion; 2. The multi-cylinder internal combustion engine according to claim 1, wherein the rate is changed stepwise. 10. The exhaust gas recirculation rate during the first combustion is substantially 55% or more, and the exhaust gas recirculation rate during the second combustion is substantially 50%. The multi-cylinder internal combustion engine according to claim 9, which is equal to or less than percent. 11. An operation region of the engine is divided into a first operation region on a low load side and a second operation region on a high load side, and the first combustion is performed in the first operation region. The multi-cylinder internal combustion engine according to claim 9, wherein the second combustion is performed in a second operation range.