JP3341685B2 - Apparatus for detecting variation in intake air amount between cylinders of an internal combustion engine - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

[0001] 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 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 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 prevent the exhaust of soot (smoke) and the exhaust of NOx from the internal combustion engine at the same time and to control the intake air of each cylinder regardless of the variation of the fuel injection amount of each cylinder. It is an object of the present invention to provide an inter-cylinder intake air amount variation detecting device capable of accurately calculating the amount variation.

[0011]

According to the first aspect of the present invention, a plurality of cylinders are provided, and as the amount of inert gas supplied into the combustion chamber increases, the amount of soot generated gradually increases. When the amount of the inert gas supplied into the combustion chamber further increases, the temperature of the fuel during combustion in the combustion chamber and the temperature of the gas around it become lower than the soot generation temperature. a cylinder-to-cylinder intake air amount variation detecting device for an internal combustion engine soot is hardly generated, when the combustion front Kisusu hardly occurs is performed by detecting the engine angular velocity in stroke including the expansion stroke of each cylinder, it is detected Further, there is provided an inter-cylinder intake air amount variation detecting apparatus for calculating a variation in intake air amount supplied to each cylinder based on the engine angular velocity.

According to the second aspect of the present invention, the cylinder angular velocity of the internal combustion engine according to the first aspect is obtained by detecting a time required for a part of a stroke including an expansion stroke of each cylinder. An inter-intake air amount variation detection device is provided.

According to the third aspect of the present invention, the engine angular velocity is obtained by detecting the time required for the entire stroke including the expansion stroke of each cylinder. An intake air amount variation detection device is provided.

According to the first to third aspects of the present invention, the amount of the inert gas supplied into the combustion chamber is smaller than the amount of the inert gas at which the generation amount of soot reaches a peak. When combustion in which little soot is generated is performed, the engine angular velocity in a stroke including the expansion stroke of each cylinder is detected, and a variation in the amount of intake air supplied to each cylinder is calculated based on the detected engine angular velocity. Here, the “stroke including the expansion stroke” refers to an expansion stroke and a compression stroke adjacent thereto. Incidentally, the combustion in which the soot is hardly generated is performed in a state where there is not much air used for the combustion. That is, when the combustion is performed in which the soot is hardly generated, it can be said that the fuel used for the combustion is slightly excessive. Therefore, even if the fuel injection amount varies somewhat, the amount of fuel used for combustion does not change, and therefore, the generated torque does not change. On the other hand, it can be said that the air used for combustion is slightly insufficient. Therefore, when the intake air amount varies, the generated torque varies significantly. Here, the variation in the generated torque of each cylinder can be obtained by detecting the engine angular velocity in a stroke including an expansion stroke of each cylinder. Therefore, as described above, claims 1 to
3. In the inter-cylinder intake air amount variation detecting device according to 3, the combustion in which the soot is hardly generated is performed.
An engine angular speed in a stroke including an expansion stroke of each cylinder is detected, and a variation in intake air amount supplied to each cylinder is calculated based on the detected engine angular speed. As a result, regardless of the variation in the fuel injection amount of each cylinder, the variation in the intake amount of each cylinder can be accurately calculated.

According to the fourth aspect of the present invention, before detecting the engine angular velocity, the variation in the fuel injection amount supplied to each cylinder is corrected. An inter-intake air amount variation detection device is provided.

According to the fifth aspect of the present invention, the first combustion, in which the soot is hardly generated, and the amount of the inert gas, in which the amount of generated soot becomes a peak, is supplied to the combustion chamber more than the amount of the inert gas. Switching means for selectively switching between the second combustion in which the amount of the inert gas is small and the second combustion in which the variation in the fuel injection amount supplied to each cylinder is corrected when the second combustion is performed. According to a fourth aspect of the present invention, there is provided an apparatus for detecting variation in intake air amount between cylinders of an internal combustion engine.

In the inter-cylinder intake air amount variation detecting apparatus according to the fourth and fifth aspects, the fuel injection amount supplied to each cylinder is detected before detecting the engine angular velocity in a stroke including an expansion stroke of each cylinder. The variation is corrected. for that reason,
The detected engine angular velocity does not include the variation due to the variation in the fuel injection amount supplied to each cylinder. As a result, the calculated variation in the intake air amount of each cylinder is not affected by the variation in the fuel injection amount of each cylinder. Therefore, it is possible to more accurately calculate the variation in the intake air amount of each cylinder.

According to the sixth aspect of the present invention, there is provided the internal combustion engine according to the first aspect, wherein a catalyst having an oxidizing function is disposed in an engine exhaust passage for oxidizing unburned hydrocarbons discharged from the combustion chamber. An apparatus for detecting variation in intake air amount between cylinders of an engine is provided.

According to the present invention, the catalyst comprises at least one of an oxidation catalyst, a three-way catalyst, and a NOx absorbent. Is provided.

According to the sixth and seventh aspects of the present invention, the unburned hydrocarbon discharged from the combustion chamber is oxidized in the engine exhaust passage. It is possible to prevent the exhaust from the internal combustion engine.

According to the present invention, there is provided an exhaust gas recirculation device for recirculating exhaust gas discharged from the combustion chamber into the engine intake passage, wherein the inert gas is supplied to the engine intake passage. 2. An apparatus for detecting variation in intake air amount between cylinders of an internal combustion engine according to claim 1, comprising a recirculated exhaust gas recirculated inside.

According to the eighth aspect of the present invention, there is provided an inter-cylinder intake amount variation detecting apparatus for an internal combustion engine, wherein recirculated exhaust gas recirculated into an engine intake passage by an exhaust gas recirculation device is used as an inert gas. The necessity of specially providing a means for supplying an inert gas from the outside into the combustion chamber can be avoided.

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 inter-cylinder intake air amount variation detecting device for an internal combustion engine according to claim 8, wherein the exhaust gas recirculation rate is changed stepwise when the switching is made.

According to the ninth aspect of the present invention, the exhaust gas recirculation is performed when switching from the first combustion to the second combustion or from the second combustion to the first combustion is performed. By changing the rate in steps,
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 tenth aspect, when the exhaust gas recirculation rate during the first combustion is substantially equal to or greater than 55%, and when the second combustion is performed. The inter-cylinder intake air amount variation detecting device for an internal combustion engine according to claim 9, wherein the exhaust gas recirculation rate of the internal combustion engine is approximately 50% or less.

In the apparatus for detecting variation in intake air amount between cylinders of an internal combustion engine according to the present invention, the exhaust gas recirculation rate during the first combustion is set to approximately 55% or more, and the second combustion is performed. Avoiding the exhaust gas recirculation rate to be set to an exhaust gas recirculation rate at which the amount of soot generation peaks by making the exhaust gas recirculation rate when it is performed approximately 50% or less. Can be.

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.
The cylinder of the internal combustion engine according to claim 9, wherein the first combustion is performed in the first operation range, and the second combustion is performed in the second operation range. An inter-intake air amount variation detection device is provided.

In the apparatus for detecting variation in intake air amount between cylinders of an internal combustion engine according to the present invention, when the first combustion can be executed,
In other words, when the temperature of the fuel and the surrounding gas during combustion in the combustion chamber can be maintained lower than the temperature at which soot is generated, the time during which the calorific value due to combustion is relatively low during engine low-load operation is limited. The first combustion is performed in the first operation region on the load side, and the second combustion is performed in the second operation region on the high load side.
Combustion. Therefore, appropriate combustion can be performed according to the operation range.

[0029]

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 an embodiment applied to a cylinder internal combustion engine. FIG.
, 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,
Reference numeral 9 denotes an exhaust valve, and reference numeral 10 denotes an exhaust port. 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 a supercharger, for example, an outlet of a compressor 16 of an exhaust turbocharger 15 via an intake duct 13 and an intercooler 14. Be linked. An inlet of the compressor 16 is connected to an air cleaner 18 via an air suction pipe 17, and a throttle valve 20 driven by a step motor 19 is provided in the air suction pipe 17.
Is arranged. 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.

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

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

The electronic control unit 40 is composed of a digital computer, and is connected to a ROM (read only memory) 42, a RAM (random access memory) 43, a CPU (microprocessor) 44, an input port 45, An output port 46 is provided. The output signal of the 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. 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. . Also,
The input port 45 has a crank angle sensor 52 that generates an output pulse every time the crankshaft rotates, for example, by 30 °.
Is connected. Further, an acceleration sensor 60 that generates an output pulse every time the camshaft rotates 360 ° is connected to the input port 45 via an AD converter 47. Based on the output signals of the crank angle sensor 52 and the acceleration sensor 60, the engine angular velocity in the stroke including the expansion stroke of each cylinder is detected. On the other hand, the output port 46 is connected to a fuel injection valve 6, a throttle valve control step motor 19, an EGR control valve control step motor 30 via a corresponding drive circuit 48.
And the fuel pump 35.

FIG. 2 shows the throttle valve 2 when the engine is operating at a low load.
The graph 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. 2) is changed by changing the opening degree and the EGR rate of 0. 7 shows an experimental example. As can be seen from FIG. 2, in this experimental example, the smaller the air-fuel ratio A / F, the higher the EGR rate. When the air-fuel ratio A / F is smaller than the stoichiometric air-fuel ratio (≒ 14.6), the EGR rate is 65% or more.

As shown in FIG. 2, when the air-fuel ratio A / F is reduced by increasing the EGR rate, the smoke is reduced when the EGR rate becomes 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. 3 (A) shows a 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, and FIG. 3 (B) 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. 3 (A) and FIG. 3 (B), in the case of FIG. 3 (B) where the amount of smoke generation is almost zero, FIG.
It can be seen that the combustion pressure is lower than in the case shown in (A).

The following can be said from the experimental results shown in FIGS. That is, first, the air-fuel ratio A / F is 1
FIG. 2 when the smoke generation amount is almost zero at 5.0 or less.
As shown in (2), the generation amount of NOx is considerably reduced. N
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. 3B where almost no soot is generated, the combustion pressure is low.
The combustion temperature inside is low.

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

The above considerations based on the experimental results shown in FIGS. 2 and 3 can be summarized as follows. 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 air-fuel ratio compression 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 is 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 is set to a temperature lower than the temperature at which the soot is generated. It needs to be suppressed. In this case, it has been found that the endothermic effect of the gas around the fuel when the fuel burns has an extremely large effect on suppressing the temperature of the fuel and the gas around the fuel.

That is, if there is only air around the fuel, the evaporated fuel immediately reacts with oxygen in the air and burns. In this case, the temperature of the air separated from the fuel does not rise so much, and only the temperature around the fuel becomes extremely high locally. That is, at this time, the air separated from the fuel hardly absorbs the heat of combustion heat of the fuel. In this case, since the combustion temperature 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. 5 shows the relationship between the EGR rate and the 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. 5, a curve A indicates that the EGR gas temperature is substantially 9
Curve B shows the case where the EGR gas is cooled by a small cooling device, and curve C shows the case where the temperature is maintained at 0 ° C.
Indicates a case where the EGR gas is not forcibly cooled.

As shown by the curve A in FIG. 5, when the EGR gas is cooled strongly, the 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.

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

FIG. 5 shows the amount of smoke generated when the engine load is relatively high. When the engine load decreases, the EGR rate at which the amount of soot peaks slightly decreases, and 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. 6 shows the mixing of EGR gas and air necessary to make the fuel during combustion and the surrounding gas 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. 6, the vertical axis indicates the total intake gas amount sucked into the combustion chamber 5, and the dashed line Y indicates the total intake gas amount that can be sucked into the combustion chamber 5 when supercharging is not performed. ing. The horizontal axis indicates the required load.

Referring to FIG. 6, the proportion of air, that is, the amount of air in the mixed gas, indicates the amount of air necessary to completely burn the injected fuel. That is, in the case shown in FIG. 6, the ratio between the air amount and the injected fuel amount is the stoichiometric air-fuel ratio. On the other hand, in FIG. 6, the ratio of the EGR gas, that is, the amount of the EGR gas in the mixed gas, is set so that when the injected fuel is burned, the temperature of the fuel and the surrounding gas is lower than the temperature at which soot is formed. The required minimum EGR gas amount is shown. This EGR gas amount is approximately 55% or more in terms of the EGR rate, 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. 6, 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 amount of heat generated when the fuel burns increases. Therefore, in order to maintain the temperature of the fuel and the surrounding gas at a temperature lower than the temperature at which soot is generated, the heat generated by the EGR gas is required. Must be increased. Therefore, as shown in FIG. 6, the EGR gas amount must be increased as the injected fuel amount increases.
That is, the EGR gas amount needs to increase as the required load increases.

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

However, as shown in FIG. 1, when the EGR gas is recirculated through the EGR passage 29 to the inlet side of the supercharger, that is, into the air suction pipe 17 of the exhaust turbocharger 15, the required load is larger than Lo. In EGR rate 5
It can be maintained at 5% or more, for example 70%, so that the temperature of the fuel and its surrounding gas can be kept below the temperature at which soot is produced. That is, if the EGR gas is recirculated so that the EGR rate in the air suction pipe 17 becomes, for example, 70%, the EGR rate of the suction gas boosted by the compressor 16 of the exhaust turbocharger 15 also becomes 70%. The temperature of the fuel and the surrounding gas can be maintained at a temperature lower than the temperature at which soot is generated, to the extent that the pressure can be increased by the compressor 16. Therefore, the operating range of the engine that can 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. 6 shows the case where fuel is burned under the stoichiometric air-fuel ratio. However, even if the air amount is smaller than the air amount shown in FIG. 6, 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. 6, that is, even if the average value of the air-fuel ratio is 17 to 18 lean, soot generation 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 the combustion temperature is suppressed to a low temperature, so that 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.

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

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

As described above, two boundaries, that is, the first boundary X (N) and the second boundary Y (N) on the load side lower 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 carried out, soot is hardly generated, but the unburned hydrocarbon is replaced by 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.

In addition to 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. 8 shows the output of the air-fuel ratio sensor 27. As shown in FIG. 8, the output current I of the air-fuel ratio sensor 27 changes according to the air-fuel ratio A / F. Therefore, the air-fuel ratio can be known from the output current I of the air-fuel ratio sensor 27.

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

In other words, in the first operation 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. 9, the EGR rate is reduced stepwise from approximately 70% to 40% or less, and the air-fuel ratio is increased stepwise. That is, the EGR 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 operation 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. 10A shows the target air-fuel ratio A / F in the first operation region I. In FIG. 10A, 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. 10A, 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. 10 (A) is preliminarily stored in a ROM 4 as a function of the required load L and the engine speed N as shown in FIG. 10 (B).
2 is stored. Also, 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)
As shown in FIG. 11B, 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.

FIG. 12A shows the target air-fuel ratio A when the second combustion, that is, ordinary 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. Also, the throttle valve 20 required to set the air-fuel ratio to the target air-fuel ratio A / F shown in FIG.
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 as shown in FIG.
As shown in FIG. 13 (B), 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 in the form of a map 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. 15, first, Step 1
At 00, it is determined whether or not a flag I indicating that the operating state of the engine is in the first operating region I is set. When the flag I is set, that is, when the operating state of the engine is in the first operating region I, step 10
The program proceeds to 1 to determine whether the required load L has become larger than the first boundary X1 (N). When L ≦ X1 (N), the routine proceeds to step 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. 11A, and the opening of the throttle valve 20 is set to the target opening ST. Next, at step 104, the target opening SE of the EGR control valve 31 is calculated from the map shown in FIG.
The opening of the EGR control valve 31 is set as the target opening SE.
Next, at step 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. 14, 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. 13B, 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.

As described above, 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 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 the open loop when low-temperature combustion is being performed, and the air-fuel ratio is controlled by the opening degree of the throttle valve 20 when the second combustion is being 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.

Next, a method for detecting a variation in the intake air amount between the cylinders of the internal combustion engine according to the present embodiment will be described. Before describing the method of detecting the variation in the intake air amount between the cylinders of the internal combustion engine according to the present embodiment, a phenomenon that is the basis of the method will be described. FIG. 16 is a graph showing the relationship between the air-fuel ratio A / F and the generated torque T when low-temperature combustion is performed. In FIG. 16, ΔA / F ₁ indicates a variation in the air-fuel ratio due to a variation in the fuel injection amount when the low-temperature combustion is being performed.
ΔT ₁ indicates the variation of the generated torque due to the variation of the fuel injection amount when the low-temperature combustion is being performed, and ΔA
/ F ₂ indicates the variation in the air-fuel ratio due to the variation in the intake air amount when the low-temperature combustion is being performed, and ΔT ₂ is the torque generated due to the variation in the fuel injection amount when the low-temperature combustion is being performed. The variation is shown.

The low temperature combustion is performed in a state where there is not much air used for combustion. In other words, when low-temperature combustion is performed, it can be said that the fuel used for combustion is slightly excessive. Therefore, even if the fuel injection amount varies, the fuel amount actually used for combustion does not change. Therefore, as shown in FIG. 16, the variation ΔA / F ₁ of the air-fuel ratio due to the variation of the fuel injection amount is relatively small. . As a result, the variation ΔT ₁ in the generated torque due to the variation in the fuel injection amount is relatively small. On the other hand, when low-temperature combustion is performed, it can be said that the air used for combustion is slightly insufficient. therefore,
When the amount of intake air varies, the amount of air actually used for combustion changes remarkably. Therefore, as shown in FIG. 16, the variation ΔA / F ₂ of the air-fuel ratio due to the variation of the amount of intake air varies with the variation of the fuel injection amount. The resulting air-fuel ratio variation ΔA / F ₁ is larger than that. As a result, the variation ΔT ₂ in the generated torque due to the variation in the intake air amount is also larger than the variation ΔT ₁ in the generated torque due to the variation in the fuel injection amount. still,
The variation in the generated torque of each cylinder can be obtained by detecting the engine angular velocity in a stroke including an expansion stroke of each cylinder.

Therefore, in the method of detecting the variation in the intake air amount between the cylinders of the internal combustion engine according to the present embodiment, when the low-temperature combustion is performed,
An engine angular velocity in a stroke including an expansion stroke of each cylinder is detected, and a variation in intake air amount supplied to each cylinder is calculated based on the detected engine angular velocity. According to the inter-cylinder intake air amount variation detection method of the present embodiment, it is possible to accurately calculate the intake air amount variation of each cylinder regardless of the fuel injection amount variation of each cylinder.

FIG. 17 is a flowchart showing a part of the method for detecting a variation in intake air amount between cylinders of an internal combustion engine according to the present embodiment. This routine is executed at the timing of the top dead center 180 ° crank angle (TDC 180 ° CA). As shown in FIG. 17, when this routine is started, first, in step 1701, it is determined whether or not low-temperature combustion is currently being performed. When the determination is NO, the variation in the intake air amount between the cylinders cannot be detected, and thus this routine ends. On the other hand, when the result is YES, in step 1702, the crank angle sensor 52 and the acceleration sensor 60 determine which cylinder is currently in the stroke including the expansion stroke, and update the cylinder number i. The internal combustion engine of the present embodiment has a first cylinder, a third cylinder, a fourth cylinder,
Since the expansion stroke is performed in the order of the cylinder number, when it is determined that the cylinder in the stroke including the current expansion stroke is the first cylinder, the cylinder number i is set to 1 (i ← 1), and the stroke including the current expansion stroke is performed. Is determined to be the third cylinder, the cylinder number i is set to 2 (i ← 2). When it is determined that the cylinder currently in the stroke including the expansion stroke is the fourth cylinder, the cylinder number i is determined. Is set to 3 (i ← 3), and when it is determined that the cylinder in the stroke including the expansion stroke is the second cylinder, the cylinder number i is set to 4 (i ← 4).

Next, at step 1703, the time TIMER required for the stroke including the expansion stroke calculated based on the output signal of the crank angle sensor 52 is set to the time Ti required for the stroke including the expansion stroke of the i-th cylinder. (Ti ← TIM
ER). Next, at 1704, the TIMER is reset, and preparations are made to measure the time required for the stroke including the expansion stroke of the next cylinder.

That is, if the time T ₁ required for the stroke including the expansion stroke of the first cylinder, for example, is obtained when this routine is executed for the first time, the next time this routine is executed, 3
Ban time T ₂ required for the process, including an expansion stroke of the cylinder is obtained, then the time T ₃ required for stroke, including an expansion stroke of the fourth cylinder when executing this routine obtained, then executes this routine at time T ₄ required for stroke, including an expansion stroke of the second cylinder is obtained.

Next, in the method for detecting variation in intake air amount between cylinders of an internal combustion engine according to the present embodiment, in steps (not shown),
Time variation coefficient Ti × 4 required for a stroke including an expansion stroke of the i-th cylinder, which indicates a degree of variation in the intake air amount of the i-th cylinder.
/ (T ₁ + T ₂ + T ₃ + T ₄ ) is calculated. When the intake amount of the i-th cylinder is larger than the average intake amount of all the cylinders, the variation coefficient is larger than 1, and when the intake amount of the i-th cylinder is smaller than the average intake amount of all the cylinders, the variation coefficient is greater than 1. Is also smaller.

In the present embodiment, the engine angular velocity in the stroke including the expansion stroke of each cylinder is obtained by detecting the time Ti required for the entire stroke including the expansion stroke of each cylinder. In the embodiment, the engine angular velocity in the stroke including the expansion stroke of each cylinder may be obtained by detecting the time required for a part of the stroke including the expansion stroke of each cylinder.

As described above, even when the fuel injection amount varies between cylinders during low-temperature combustion, the variation in the generated torque of each cylinder is relatively small, but the variation in the intake amount of each cylinder is calculated more accurately. In another embodiment, the second combustion (combustion according to the conventional combustion method) is performed before the engine angular velocity in the stroke including the expansion stroke of each cylinder is detected at the time of low-temperature combustion. May be corrected. According to this embodiment, since the variation in the intake air amount of each cylinder calculated at the time of low-temperature combustion is not affected by the variation in the fuel injection amount of each cylinder, the variation of the fuel injection amount of each cylinder is not corrected. Variations in the intake air amount of each cylinder can be calculated more accurately than in the case.

Instead of this embodiment, that is, instead of correcting the variation in the fuel injection amount of each cylinder during the second combustion, a fuel injection valve having a small variation in the fuel injection amount between the cylinders is used. It is also possible.

[0101]

According to the first to third aspects of the present invention, soot (smoke) is discharged from the internal combustion engine and NOx
Can be simultaneously calculated and the variation in the intake air amount in each cylinder can be accurately calculated regardless of the variation in the fuel injection amount in each cylinder.

According to the fourth and fifth aspects of the present invention, without being affected by variations in the fuel injection amount of each cylinder,
Variations in the intake air amount of each cylinder can be calculated more accurately.

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

According to the eighth aspect of the present invention, it is possible to avoid the necessity of specially providing a means for supplying an inert gas from the outside into the combustion chamber.

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 a diagram showing amounts of smoke and NOx generated;

FIG. 3 is a diagram showing a combustion pressure.

FIG. 4 is a diagram showing fuel molecules.

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

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

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

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

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

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

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

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

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

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

FIG. 15 is a flowchart for controlling operation of the engine.

FIG. 16 is a graph showing a relationship between an air-fuel ratio A / F and a generated torque T when low-temperature combustion is performed.

FIG. 17 is a flowchart showing a part of a method for detecting a variation in intake air amount between cylinders of an internal combustion engine.

[Explanation of symbols]

 5 Combustion chamber 20 Throttle valve 29 EGR passage 31 EGR control valve 52 Crank angle sensor 60 Acceleration sensor

──────────────────────────────────────────────────の Continued on the front page (51) Int.Cl. ⁷ Identification code FI F01N 3/24 F01N 3/24 BES F02D 21/08 301 F02D 21/08 301H 41/40 41/40 N F02M 25/07 550 F02M 25/07 550F 550P 570 570D (72) Inventor Yoshi ▲ Saki ▼ Koji 1 Toyota Town, Toyota City, Aichi Prefecture Inside Toyota Motor Co., Ltd. (72) Inventor Hiroki Murata 1 Toyota Town, Toyota City, Aichi Prefecture Toyota Motor (56) References JP-A-7-4287 (JP, A) JP-A 8-177654 (JP, A) JP-A 8-86251 (JP, A) JP-A 9-287527 (JP) JP-A-9-287528 (JP, A) JP-A-9-217658 (JP, A) JP-A-9-79092 (JP, A) JP-A-9-14026 (JP, A) (58) investigated Field (Int.Cl. ^7, DB name) F02D 41/00 - 45/00 395

Claims (11)

(57) [Claims] 1. A fuel supply system comprising a plurality of cylinders, the amount of generated soot gradually increases as the amount of inert gas supplied into the combustion chamber increases, reaches a peak, and the amount of inert gas supplied into the combustion chamber increases. As the amount of the active gas is further increased, the temperature of the fuel during combustion in the combustion chamber and the temperature of the gas around it become lower than the temperature at which soot is generated, and soot generation hardly occurs. an apparatus, prior to when the Kisusu is hardly generated combustion is performed, and detecting the engine angular velocity in stroke including the expansion stroke of each cylinder, intake air quantity supplied to the cylinders based on the detected engine angular velocity A variation detection device for calculating the intake air amount between cylinders of an internal combustion engine, which calculates the variation of the intake air amount. 2. The apparatus according to claim 1, wherein the engine angular velocity is obtained by detecting a time required for a part of a stroke including an expansion stroke of each cylinder. 3. The inter-cylinder intake air amount variation detecting device according to claim 1, wherein the engine angular velocity is obtained by detecting a time required for the entire stroke including an expansion stroke of each cylinder. 4. The inter-cylinder intake air amount variation detecting device according to claim 1, wherein a variation in a fuel injection amount supplied to each cylinder is corrected before the engine angular speed is detected. 5. The first combustion, in which the soot is hardly generated, and the second combustion, in which the amount of inert gas supplied into the combustion chamber is smaller than the amount of inert gas at which the amount of generated soot becomes a peak. 5. A switching means for selectively switching between the second combustion and the second combustion, wherein a variation in a fuel injection amount supplied to each cylinder is corrected when the second combustion is performed.
3. The apparatus for detecting a variation in intake air amount between cylinders of an internal combustion engine according to claim 1. 6. An apparatus according to claim 1, wherein a catalyst having an oxidizing function is disposed in an engine exhaust passage for oxidizing unburned hydrocarbons discharged from the combustion chamber. . 7. The catalyst according to claim 1, wherein the catalyst is an oxidation catalyst, a three-way catalyst or NO.
7. The apparatus according to claim 6, wherein the apparatus comprises at least one of x absorbents. 8. An exhaust gas recirculation device for recirculating exhaust gas discharged from the combustion chamber into an engine intake passage, wherein the inert gas is recirculated into the engine intake passage. 2. The inter-cylinder intake air amount variation detecting apparatus according to claim 1, wherein the apparatus comprises gas. 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; 9. The apparatus according to claim 8, 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 inter-cylinder intake air amount variation detecting apparatus according to claim 9, wherein the percentage is not more than a percentage. 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 apparatus according to claim 9, wherein the second combustion is performed in a second operation range.