JP3409722B2 - Exhaust gas recirculation control valve - Google Patents

Description

Detailed Description of the Invention

[0001]

TECHNICAL FIELD The present invention relates to an exhaust gas recirculation amount control valve.

[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 into the engine intake passage through the 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 NOx generation amount decreases, and therefore, the higher the EGR rate, the lower the NOx generation amount.

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

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

As described above, in the past, it was considered that the maximum allowable limit exists for the EGR rate.
The R rate is NO within the range that does not exceed this maximum allowable limit.
It was specified that the amount of x and smoke generated should be as small as possible. However, in this way EGR
Even if the rate is set so that the amount of NOx and smoke produced is as small as possible, there is a limit to the reduction in the amount of NOx and smoke produced, and in reality, a considerable amount of N 2 is still generated.
The present situation is that Ox and smoke are generated.

FIG. 18 is a partial sectional side view of an EGR control valve provided in a conventional EGR passage. In FIG. 18, 1000 is an EGR control valve, 1001 is an EGR passage,
Reference numeral 1002 is a valve element, and 1003 is an EGR control valve drive unit. As shown in FIG. 18, in the conventional EGR control valve, EG
In order to control the amount of R gas, the valve body 1002 strokes in the direction of the EGR gas flow to change the flow passage area of EGR gas. Such a stroke type EG
The merit of adopting the R control valve is that, for example, due to the momentum that the valve body 1002 abuts during the valve closing operation, the valve body 10
There is a point that the deposit attached to 02 can be separated from the valve body 1002. An example of such an EGR control valve is disclosed in Japanese Patent Laid-Open No. 8-14114.

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

[0008] That is, as a result of repeated experimental research, it was found that when the temperature of the fuel and the gas around it during combustion in the combustion chamber were below a certain temperature, the growth of hydrocarbons stopped in the middle of the stage before reaching soot. However, if the temperature of the fuel and the gas around it rises above a certain temperature, the hydrocarbons will suddenly grow to soot. In this case, the temperature of the fuel and its surrounding gas is greatly affected by the endothermic action of the gas around the fuel when the fuel burns, and the endothermic amount of the gas around the fuel is adjusted according to the amount of heat generated during fuel combustion. Thus, the temperature of the fuel and the gas around it can be controlled.

Therefore, if the temperature of the fuel and the gas around it during combustion in the combustion chamber is suppressed below the temperature at which the growth of hydrocarbons stops midway, soot will not be generated, and the fuel and the surroundings during combustion in the combustion chamber will be eliminated. It is possible to control the gas temperature in the range below the temperature at which the growth of hydrocarbons stops halfway by adjusting the heat absorption amount of the gas around the fuel. On the other hand, hydrocarbons whose growth has stopped before reaching soot can be easily purified by post-treatment using an oxidation catalyst or the like. This is the basic idea of the new combustion system.

[0010]

In such a new combustion system, the EGR rate is set to a considerably large value as described above. Therefore, in order to make the EGR rate considerably large, it is necessary to make the maximum flow passage area of the EGR gas considerably large. Further, in order to make the EGR rate considerably large, the intake air amount is considerably narrowed. When the intake air amount is throttled, a negative pressure is generated in the engine intake passage downstream of the portion where the intake air amount is throttled. Therefore, the EG is controlled so that the opening degree of the EGR control valve does not change even if it is affected by the negative pressure.
It is necessary to configure the R control valve.

Therefore, according to the present invention, while preventing soot and NOx from being discharged from the internal combustion engine at the same time, even when a negative pressure is generated in the engine intake passage, the EGR control valve To provide an exhaust gas recirculation amount control valve capable of maintaining a desired opening amount and increasing the maximum flow passage area of the recirculated exhaust gas without increasing the outer shape of the exhaust gas recirculation passage. With the goal.

[0012]

According to the present invention, the recirculation exhaust gas is provided between the engine exhaust passage and the engine intake passage.
Recirculation exhaust gas that is connected to the combustion chamber and is supplied to the combustion chamber.
As the amount of gas increases, the amount of soot generated gradually increases
The amount of recirculated exhaust gas that reaches the
As it further increases, the fuel and
And the temperature of the gas around it is lower than the soot formation temperature
An exhaust gas recirculation amount control valve provided for controlling the amount of recirculated exhaust gas in the exhaust gas recirculation passage of an internal combustion engine, in which the valve body rotates to recirculate exhaust gas. An exhaust gas recirculation amount control valve is provided, in which a flow passage area of gas is changed and a rotation axis of the valve body is arranged at a center of the valve body in a direction perpendicular to a flow direction of recirculated exhaust gas. It

According to a second aspect of the invention, there is provided the exhaust gas recirculation amount control valve according to the first aspect, wherein the exhaust gas recirculation amount control valve is a butterfly type.

According to the invention described in claim 3, there is provided the exhaust gas recirculation amount control valve according to claim 1, wherein the exhaust gas recirculation amount control valve is a rotary type.

In the exhaust gas recirculation amount control valve according to any one of claims 1 to 3, the amount of the recirculation exhaust gas supplied to the combustion chamber is larger than the amount of the recirculation exhaust gas at which the amount of generated soot peaks. Since combustion in which soot is hardly generated is performed at a low temperature, it is possible to simultaneously prevent soot and NOx from being discharged from the internal combustion engine. Further, the rotary shaft of the valve body of the exhaust gas recirculation amount control valve is arranged perpendicular to the flow direction of the recirculated exhaust gas and at the center of the valve body. Therefore, the force received by the negative pressure on the valve body on one side with respect to the rotation axis of the valve body and the force received on the other side by the negative pressure with respect to the rotation axis of the valve body are offset. As a result, the exhaust gas recirculation amount control valve is not affected by the negative pressure even under the condition that a negative pressure is generated in the engine intake passage. Maintained in degrees. Further, since the rotary shaft of the valve body of the exhaust gas recirculation amount control valve is arranged perpendicular to the flow direction of the recirculated exhaust gas and at the center of the valve body, the outer shape of the exhaust gas recirculation passage is The maximum flow passage area of the recirculated exhaust gas can be made larger than the case of the conventional stroke type EGR control valve shown in FIG. 18 without increasing the size.

According to the invention described in claim 4, in the internal combustion engine, the first combustion, which is a combustion in which the soot is hardly generated, and the amount of the recirculated exhaust gas, in which the soot generation amount reaches a peak, Switching means for selectively switching between the second combustion in which the amount of recirculated exhaust gas supplied to the combustion chamber is small, and the first combustion to the second combustion or the second combustion. The exhaust gas recirculation amount control valve according to claim 1, wherein the exhaust gas recirculation rate is changed stepwise when switching to the first combustion.

In the exhaust gas recirculation amount control valve according to the fourth aspect, the exhaust gas recirculation rate is stepped 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 can be prevented from being set to the exhaust gas recirculation rate at which the amount of soot generated reaches a peak by changing the state.

According to the invention described in claim 5, the first
Exhaust gas recirculation rate is almost 5 when combustion of
The exhaust gas recirculation amount control valve according to claim 4, wherein the exhaust gas recirculation rate is 5% or more and the exhaust gas recirculation rate when the second combustion is performed is approximately 50% or less.

In the exhaust gas recirculation amount control valve according to the fifth aspect, the exhaust gas recirculation rate during the first combustion is set to about 55% or more and the second combustion is performed. By setting the exhaust gas recirculation rate when the exhaust gas recirculates to approximately 50% or less, it is possible to prevent the exhaust gas recirculation rate from being set to the exhaust gas recirculation rate at which the amount of soot generated peaks.

According to the invention as set forth in claim 6, the operating region of the engine is divided into a first operating region on the low load side and a second operating region on the high load side, and in the first operating region, The first
The exhaust gas recirculation amount control valve according to claim 4, wherein the exhaust gas recirculation amount control valve is configured to perform the second combustion in the second operation region.

In the exhaust gas recirculation amount control valve according to the sixth aspect, the temperature of the fuel and the gas around it when the first combustion can be executed, that is, the combustion temperature in the combustion chamber is lower than the soot generation temperature. Since the time that can be kept low is limited to the engine low load operation in which the calorific value due to combustion is relatively small, the first combustion is performed in the low load side first operation region and the high load side second The second combustion is performed in the operating region of. Therefore, appropriate combustion can be executed according to the operating region.

According to a seventh aspect of the present invention, the internal combustion engine comprises a catalyst having an oxidizing function for oxidizing unburned hydrocarbons discharged from the combustion chamber, in the engine exhaust passage. An exhaust gas recirculation amount control valve as described in 1. is provided.

According to the invention described in claim 8, there is provided the exhaust gas recirculation amount control valve according to claim 7, wherein the catalyst comprises at least one of an oxidation catalyst, a three-way catalyst and a NOx absorbent. It

In the exhaust gas recirculation amount control valve according to the seventh and eighth aspects, since unburned hydrocarbons discharged from the combustion chamber are oxidized in the engine exhaust passage, unburned hydrocarbons are discharged from the internal combustion engine. It can be prevented from being discharged.

[0025]

BEST MODE FOR CARRYING OUT THE INVENTION Embodiments of the present invention will be described below with reference to the accompanying drawings.

FIG. 1 shows a first embodiment in which the present invention is applied to a 4-stroke compression ignition type internal combustion engine. Referring to FIG. 1, 1 is an engine body, 2 is a cylinder block, 3
Is a cylinder head, 4 is a piston, 5 is a combustion chamber, 6 is an electrically controlled fuel injection valve, 7 is an intake valve, 8 is an intake port, 9
Indicates an exhaust valve, and 10 indicates an exhaust port, respectively. 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 portion of a compressor 16 of an exhaust turbocharger 15 via an intake duct 13 and an intercooler 14. .
The inlet of the compressor 16 is connected to an air cleaner 18 via an air suction pipe 17, and a throttle valve 20 driven by a step motor 19 is arranged in the air suction pipe 17. In addition, the air suction pipe 1 upstream of the throttle valve 20
A mass flow rate detector 21 for detecting the mass flow rate of the intake air is arranged inside 7.

On the other hand, the exhaust port 10 is connected to the exhaust manifold 2
Exhaust turbine 2 of exhaust turbocharger 15 via 2
3 and the outlet of the exhaust turbine 23 is connected via an exhaust pipe 24 to a catalytic converter 26 containing a catalyst 25 having an oxidizing function. Exhaust manifold 22
An air-fuel ratio sensor 27 is arranged inside.

The exhaust pipe 28 connected to the outlet of the catalytic converter 26 and the air intake pipe 17 downstream of the throttle valve 20 are connected to each other via an exhaust gas recirculation (hereinafter referred to as EGR) passage 29, and the EGR passage A butterfly type EGR control valve 3 driven by a step motor 30
1 is placed. An intercooler 32 for cooling the EGR gas flowing in the EGR passage 29 is arranged in the EGR passage 29. In the embodiment shown in FIG. 1, 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 electrically controlled variable discharge fuel pump 35, and the fuel supplied into the common rail 34 is supplied to the fuel injection valve 6 via each fuel supply pipe 33. A fuel pressure sensor 36 for detecting a fuel pressure in the common rail 34 is attached to the common rail 34, and a fuel pump 35 is arranged so that the fuel pressure in the common rail 34 becomes a target fuel pressure based on an output signal of the fuel pressure sensor 36. Is controlled.

The electronic control unit 40 is composed of a digital computer and has a ROM (Read Only Memory) 42, a RAM (Random Access Memory) 43, a CPU (Microprocessor) 44, an input port 45 and an input port 45 which are connected to each other by a bidirectional bus 41. The output port 46 is provided. The output signal of the mass flow rate 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 ports via the corresponding AD converter 47. 45 is input. A load sensor 51 that generates an output voltage proportional to the depression amount L of the accelerator pedal 50 is connected to the accelerator pedal 50, and the output voltage of the load sensor 51 is input to the input port 45 via the corresponding AD converter 47. . Also,
The input port 45 has a crank angle sensor 52 that generates an output pulse each time the crankshaft rotates, for example, 30 °.
Are connected. On the other hand, the output port 46 is connected via the corresponding drive circuit 48 to the fuel injection valve 6, the throttle valve control step motor 19, and the EGR control valve control step motor 3.
0 and the fuel pump 35.

FIG. 2 is a detailed view of the butterfly type EGR control valve 31 shown in FIG. Specifically, FIG. 2A shows EG
Butterfly type EG when fully closed as seen from the direction of R gas flow
2B is a partial cross-sectional side view of the R control valve, FIG. 2B is a partial cross-sectional side view of the butterfly-type EGR control valve at the time of full opening as seen from the direction of the EGR gas flow, and FIG. 2C is a butterfly-type EGR at the time of full opening. It is a perspective view of a control valve. As shown in FIG. 2, in the EGR control valve 31 of the present embodiment, the rotary shaft 61 of the valve element is
It is arranged perpendicularly to the flow direction of the GR gas, and at the same time, it is arranged in the center of the valve body 60. That is, the valve body 60 is line-symmetrical with the rotating shaft 61 of the valve body.
The flow passage area of the EGR gas is changed by rotating the valve body 60 around the rotary shaft 61 of the valve body.

As a result of the EGR control valve 31 being configured as described above, the upper portion (FIG. 2A) of the valve body 60 with respect to the rotary shaft 61 of the valve body is received by the negative pressure in the air suction pipe 17. The force and the force received by the negative pressure on the lower portion of the valve body 60 (FIG. 2A) cancel each other out. Therefore, the air suction pipe 17
Even under the condition that a negative pressure is generated in the air suction pipe 17 as in the case of the present embodiment in which the throttle valve 20 is arranged inside,
The EGR control valve 31 is not affected by the negative pressure and therefore E
The opening degree of the GR control valve 31 can be maintained at a desired opening degree.

As a result of the EGR control valve 31 being configured as described above, as shown in FIG. 2B, the EGR passage width W1 is made to match the maximum passage width W2 of EGR gas. On the other hand, in the case of the stroke type EGR control valve as shown in FIG. 18, the EGR passage width W1 ′ is the maximum flow passage width W.
It will be bigger than 2 '. That is, according to this embodiment, the EG passage width W1 is not increased and the EG passage width W1 is increased.
The maximum flow path width W2 of R gas is the maximum flow path width W of the conventional case.
It can be greater than 2 '.

FIG. 3 is a rotary type EG of the second embodiment.
It is a detailed view of an R control valve 531. For details, see FIG.
3 is a partial cross-sectional side view of the rotary type EGR control valve at the time of fully closing seen from the direction perpendicular to the direction of EGR gas flow, FIG.
(B) is a partial cross-sectional side view of the rotary EGR control valve when it is fully opened as seen from a direction perpendicular to the direction of EGR gas flow,
FIG. 2C is a perspective view of the rotary EGR control valve. In the rotary type EGR control valve 531 shown in FIG.
Similar to the butterfly type EGR control valve 31 shown in FIG.
The rotary shaft 561 of the valve body is arranged perpendicularly to the direction of EGR gas flow, and at the same time, it is arranged in the center of the valve body 560. That is, when fully closed, the valve element 560 is line-symmetrical when viewed from the direction of EGR gas flow by the rotary shaft 561 of the valve element. The passage area of the EGR gas is changed by rotating the valve body 560 around the rotary shaft 561 of the valve body.

As a result of the EGR control valve 531 being configured as described above, the force exerted by the negative pressure in the air suction pipe on the upper portion (FIG. 3A) of the valve body 560 with respect to the rotary shaft 561 of the valve body is increased. , The force received by the negative pressure on the lower portion of the valve element 560 (FIG. 3A) is offset. Therefore, even under the condition that a negative pressure is generated in the air suction pipe, the EGR control valve 531 is not affected by the negative pressure, and therefore the opening degree of the EGR control valve 531 can be maintained at a desired opening degree.

As a result of the EGR control valve 531 being configured as described above, the EGR passage width is made to match the maximum EGR gas passage width, as in the first embodiment shown in FIG. That is, also in the present embodiment, the maximum flow passage width of EGR gas can be made larger than the maximum flow passage width in the conventional case without increasing the EGR passage width.

Returning to the description of the first embodiment, FIG. 4 shows that the air-fuel ratio A / F (horizontal axis in FIG. 4 ) is changed by changing the opening of the throttle valve 20 and the EGR rate during engine low load operation. Output torque change and smoke,
The example of an experiment showing change of the amount of discharge of HC, CO, and NOx is shown. As can be seen from FIG. 4, in this experimental example, the air-fuel ratio A
As / F decreases, the EGR rate increases, and when the theoretical air-fuel ratio (≈14.6) or less, the EGR rate is 65% or more.

As shown in FIG. 4, when the air-fuel ratio A / F is reduced by increasing the EGR rate, the EGR rate becomes around 40%, and when the air-fuel ratio A / F becomes about 30, smoke is generated. The amount of generation begins to increase. Next, when the EGR rate is further increased and the air-fuel ratio A / F is made smaller, the amount of smoke generated sharply increases and reaches a peak. Next, when the EGR rate is further increased and the air-fuel ratio A / F is reduced, the smoke sharply decreases this time, the EGR rate is increased to 65% or more, and the smoke becomes almost zero when the air-fuel ratio A / F is around 15.0. . That is, soot is hardly generated. At this time, the output torque of the engine slightly decreases, and N
The amount of Ox generated is considerably low. On the other hand, at this time, HC,
The amount of CO generated starts to increase.

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

From the experimental results shown in FIGS. 4 and 5, the following can be said. That is, first of all, the air-fuel ratio A / F is 1
When the amount of smoke generated is 5.0 or less and the amount of smoke is almost zero, FIG.
As shown in (3), the amount of NOx generated is considerably reduced. N
A decrease in the amount of generated Ox means that the combustion temperature in the combustion chamber 5 has decreased, and therefore, it can be said that the combustion temperature in the combustion chamber 5 is low when soot is hardly generated. . The same can be said from FIG. That is, the combustion pressure is low in the state shown in FIG. 5 (B) in which almost no soot is generated.
The combustion temperature inside is low.

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

Summarizing these considerations based on the experimental results shown in FIGS. 4 and 5, when the combustion temperature in the combustion chamber 5 is low, the amount of soot generated becomes almost zero, and at this time, the soot precursor or the soot precursor The hydrocarbons in this state are discharged from the combustion chamber 5. As a result of further detailed experimental research on this, when the temperature of the fuel and the gas around it in the combustion chamber 5 is below a certain temperature, the soot growth process stops halfway, that is, the soot is generated. It was found that soot was not generated at all and soot was generated when the temperature of the fuel and its surroundings in the combustion chamber 5 reached a certain temperature or higher.

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

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

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

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

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

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

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

As shown by the curve A in FIG. 7, when the EGR gas is strongly cooled, the soot generation amount peaks when the EGR rate is slightly lower than 50%, and in this case, the EGR rate is almost 55. Almost no soot is generated if the percentage is exceeded.

On the other hand, as shown by the curve B in FIG.
When the R gas is slightly cooled, the 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, soot is hardly generated.

Further, as shown by the curve C in FIG. 7, EG
When the R gas is not forcibly cooled, the EGR rate is 5
The soot generation amount peaks near 5%, and in this case, if the EGR rate is set to approximately 70% or more, soot is hardly generated.

FIG. 7 shows the amount of smoke produced when the engine load is relatively high. When the engine load becomes small, the EGR rate at which the amount of soot produced reaches a peak drops slightly, and soot is almost produced. The lower limit of the EGR rate that disappears also decreases slightly. Thus, the lower limit of the EGR rate at which soot is hardly generated changes depending on the cooling degree of EGR gas and the engine load.

FIG. 8 shows a mixture of EGR gas and air required to bring the temperature of the fuel and its surrounding gas at the time of combustion to a temperature lower than the temperature at which soot is produced when EGR gas is used as the inert gas. The amount of gas, the ratio of air in this mixed gas amount, and the ratio of EGR gas in this mixed gas are shown. In addition, in FIG. 8, the vertical axis represents the total intake gas amount sucked into the combustion chamber 5, and the chain line Y represents the total intake gas amount that can be sucked into the combustion chamber 5 when supercharging is not performed. ing. The horizontal axis shows the required load.

Referring to FIG. 8, the ratio of air, that is, the amount of air in the mixed gas indicates the amount of air required to completely burn the injected fuel. That is, in the case shown in FIG. 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 the temperature of the fuel and its surrounding gas becomes lower than the temperature at which soot is formed when the injected fuel is burned. The minimum required EGR gas amount is shown. The EGR gas amount is approximately 55% or more when expressed by the EGR rate, and is 70% or more in the embodiment shown in FIG. That is, the total amount of intake gas sucked into the combustion chamber 5 is shown by the solid line X in FIG. 8, and the ratio of the amount of air to the amount of EGR gas in this total intake gas amount X is set as shown in FIG. The temperature of the fuel and the gas around it is lower than the temperature at which soot is produced, and thus no soot is generated. Also, NOx at this time
The amount of NOx generated is around 10 p.pm or less, so the amount of NOx generated is extremely small.

Since the amount of heat generated when the fuel burns increases as the fuel injection amount increases, in order to maintain the temperature of the fuel and the gas around it at a temperature lower than the temperature at which soot is generated, heat generated by the EGR gas is used. The amount of absorption must be increased. Therefore, as shown in FIG. 8, the EGR gas amount must be increased as the injected fuel amount is increased.
That is, the EGR gas amount needs to increase as the required load increases.

By the way, when the supercharging is not performed, the upper limit of the total intake gas amount X drawn into the combustion chamber 5 is Y. Therefore, in FIG. 8, the required load is greater than Lo in the required load region. The air-fuel ratio cannot be maintained at the stoichiometric air-fuel ratio unless the EGR gas ratio is reduced as it becomes larger. In other words, when supercharging is not performed and the air-fuel ratio is maintained at the stoichiometric air-fuel ratio in a region where the required load is larger than Lo, the EGR rate decreases as the required load increases, thus In the region where the required load is larger than Lo, the temperature of the fuel and the gas around it cannot be maintained below the temperature at which soot is generated.

However, as shown in FIG. 1, when the EGR gas is recirculated to the inlet side of the supercharger, that is, the air suction pipe 17 of the exhaust turbocharger 15 through the EGR passage 29, the required load is larger than Lo. EGR rate at 5
It can be maintained above 5 percent, for example 70 percent, thus maintaining the fuel and surrounding gas temperatures below the temperature at which soot is produced. That is, if the EGR gas is recirculated so that the EGR rate in the air suction pipe 17 becomes, for example, 70%, the EGR rate of the intake gas boosted by the compressor 16 of the exhaust turbocharger 15 also becomes 70%. The temperature of the fuel and the gas around it can be maintained below the temperature at which soot is produced, up to the limit that can be boosted by the compressor 16. Therefore, the operating range of the engine capable of producing the low temperature combustion can be expanded. 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 in which the GR control valve 31 is fully opened
Is closed a little.

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 made smaller than that shown in FIG. 8, that is, the air-fuel ratio is made rich. Even so, the amount of NOx generated is 10 p.p. while preventing the generation of soot.
It can be around m or less, and even if the air amount is made larger than that shown in FIG. 8, that is, even if the average value of the air-fuel ratio is lean from 17 to 18, soot generation is prevented. Meanwhile, the amount of NOx generated 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 excessive fuel does not grow to soot, and soot is generated. There is no. Further, 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 becomes high, but in the present invention the combustion temperature is suppressed to a low temperature, soot Not generated at all. Furthermore, NOx
Also produces only a very small amount.

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 amount of NOx generated is extremely small. Therefore, considering the improvement of the fuel consumption rate, it can be said that it is preferable to make the average air-fuel ratio lean at this time.

By the way, the temperature of the fuel during combustion in the combustion chamber and the temperature of the gas around it can be suppressed below the temperature at which the growth of hydrocarbons stops halfway only when the engine is operating at low load, where the calorific value of combustion is relatively small. To be Therefore, in the embodiment according to the present invention, during low load operation in the engine, the temperature of the fuel and the gas around it during combustion is suppressed below the temperature at which the growth of hydrocarbons stops halfway, and the first combustion, that is, low temperature combustion is performed. In addition, the second combustion, that is, the combustion normally performed from the conventional one, is performed during the engine high load operation. It should be noted that here, the first combustion, that is, low temperature combustion, as is clear from the above description, the amount of inert gas in the combustion chamber is larger than the amount of inert gas at which the amount of soot generated peaks, and soot is almost generated. The second combustion, that is, the combustion that is normally performed in the past, is the combustion that does not have the amount of soot generated and the amount of the inert gas in the combustion chamber is smaller than the amount of the inert gas that does not reach the peak. Say that.

FIG. 9 shows the first operating region I where the first combustion, that is, the low temperature combustion is performed, and the second operating region II where the second combustion, that is, the combustion by the conventional combustion method is performed. There is. In FIG. 9, the vertical axis L represents the depression amount of the accelerator pedal 50, that is, the required load, and the horizontal axis N represents the engine speed. Further, in FIG. 7, X (N) is the first
Shows the first boundary between the operating region I and the second operating region II, where Y (N) is the first operating region I and the second operating region.
The second boundary with II is shown. The determination of the change of the operating region from the first operating region I to the second operating region II is made based on the first boundary X (N), and the change from the second operating region II to the first operating region II is performed.
The determination of the change of the operating range to the operating range I of the second boundary Y
It is performed based on (N).

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

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

By the way, when the engine operating region is in the first operating region I and low temperature combustion is performed, soot is scarcely generated, and instead, unburned hydrocarbons are in the state of the precursor of soot or the state before it. It is discharged from the combustion chamber 5 in shape. At this time, the unburned hydrocarbons discharged from the combustion chamber 5 are satisfactorily oxidized 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. The NOx absorbent is NOx when the average air-fuel ratio in the combustion chamber 5 is lean.
And has a 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 potassium K, sodium N
a, at least one selected from alkali metals such as lithium Li and cesium Cs, alkaline earths such as barium Ba and calcium Ca, rare earths such as lanthanum La and yttrium Y, and a noble metal such as platinum Pt. Is carried.

Not only oxidation catalysts, but also three-way catalysts and NO
The x-absorbent also has an oxidizing function, so that the three-way catalyst and the 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 operating region I
The operation control in the second operation area II will be briefly 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 operating region I where the required load L is low, the opening degree of the throttle valve 20 is gradually increased from near full closing to about 2/3 opening degree as the required load L is increased. 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. 11, E in the first operating region I
The GR rate is about 70%, and the air-fuel ratio is a slightly lean lean air-fuel ratio.

In other words, in the first operating region I, EGR
The opening of the throttle valve 20 and the opening of the EGR control valve 31 are controlled so that the ratio becomes approximately 70% and the air-fuel ratio becomes a slightly lean air-fuel ratio. Further, in the first operation region I, fuel injection is performed before the compression top dead center TDC. In this case, the injection start timing θS becomes late as the required load L becomes high, and the injection completion timing θE also becomes late as the injection start timing θS becomes late.

When the engine is idling, the throttle valve 2
0 is closed until it is almost completely closed. At this time, the EGR control valve 31
Is closed to near full closure. When the throttle valve 20 is closed to near full closure, the pressure in the combustion chamber 5 at the beginning of compression becomes low, and the compression pressure becomes small. When the compression pressure becomes small, the compression work by the piston 4 becomes small, so that the vibration of the engine body 1 becomes small. That is, in order to suppress the vibration of the engine body 1 during idle operation, the throttle valve 2
The valve is closed until 0 is almost fully closed.

On the other hand, the operating region of the engine is the first operating region I.
When changing from the second operating range II to the second operating range II, the opening degree of the throttle valve 20 is increased stepwise from about 2/3 opening degree toward the full opening direction. At this time, in the example shown in FIG. 9, the EGR rate is reduced stepwise from approximately 70% to 40% or less, and the air-fuel ratio is increased stepwise. That is, the EGR that produces a large amount of smoke with an EGR rate
The operating range of the engine is the first because the rate range (Fig. 7) is exceeded.
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 kept fully open except for a part, and the opening degree of the EGR control valve 31 is gradually reduced as the required load L increases. Further, in this operating region II, the EGR rate becomes lower as the required load L becomes higher, and the air-fuel ratio becomes smaller as the required load L becomes higher. However, the air-fuel ratio is a lean air-fuel ratio even if the required load L becomes high. Further, in the second operation region II, the injection start timing θS is set near the compression top dead center TDC.

FIG. 12A shows the target air-fuel ratio A / F in the first operating region I. In FIG. 12A, A / F = 15.5, A / F = 16, A / F = 17,
Each of the curves indicated by A / F = 18 has a target air-fuel ratio of 1
The values are 5.5, 16, 17, and 18, and the air-fuel ratio between the curves is determined by proportional distribution. 12
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 is leaner as the required load L is lower.

That is, the lower the required load L, the smaller the amount of heat generated by combustion. Therefore, as the required load L decreases, low temperature combustion can be performed even if the EGR rate is decreased.
When the EGR rate is reduced, the air-fuel ratio increases, and therefore, as shown in FIG. 12 (A), the target air-fuel ratio A / F increases as the required load L decreases. Target air-fuel ratio A /
The fuel consumption rate increases as F increases, and therefore, 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. 12 (A) is a function of the required load L and the engine speed N as shown in FIG.
It is stored in 2. In addition, the throttle valve 2 necessary for setting the air-fuel ratio to the target air-fuel ratio A / F shown in FIG.
The target opening degree ST of 0 is stored in advance in the ROM 42 in the form of a map as a function of the required load L and the engine speed N as shown in FIG.
EGR required to achieve the target air-fuel ratio A / F shown in (A)
The target opening degree SE of the control valve 31 is stored in advance in the ROM 42 in the form of a map as a function of the required load L and the engine speed N as shown in FIG. 13 (B).

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 is shown. In addition, in FIG.
= 24, A / F = 35, A / F = 45, and A / F = 60, the respective curves indicated by the target air-fuel ratios of 24, 35, 45, 6 respectively.
0 is shown. Target air-fuel ratio A shown in FIG.
/ F is a map of the ROM 42 in advance as a function of the required load L and the engine speed N as shown in FIG.
It is stored in. In addition, the throttle valve 20 necessary for setting the air-fuel ratio to the target air-fuel ratio A / F shown in FIG.
The target opening degree ST of 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. 15 (A), and the air-fuel ratio is shown in FIG. 14 (A).
The target opening degree SE of the EGR control valve 31 required to achieve the target air-fuel ratio A / F shown in Fig. 15 is shown 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. 15B. It is stored in the ROM 42.

Further, 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. This fuel injection amount Q is stored in advance in the ROM 42 in the form of a map as a function of the required load L and the engine speed N as shown in FIG.

Next, the operation control of this embodiment will be described with reference to FIG. Referring to FIG. 17, first, at step 100, it is judged if the flag I indicating that the engine operating state is the first operating region I is set or not. When flag I is set,
That is, when the operating state of the engine is in the first operating region I, the routine proceeds to step 101, where the required load L is the first boundary X
It is determined whether or not it is larger than (N). L ≦ X
In the case of (N), the routine proceeds to step 103, where low temperature combustion is performed.

That is, at step 103, the target opening degree ST of the throttle valve 20 is calculated from the map shown in FIG. 13 (A), and the opening degree of the throttle valve 20 is made this target opening degree ST. Next, at step 104, the target opening degree SE of the EGR control valve 31 is calculated from the map shown in FIG. 13 (B),
The opening degree of the EGR control valve 31 is set to this target opening degree SE.
Next, at step 105, the mass flow rate Ga of intake air (hereinafter, simply referred to as intake air amount) Ga detected by the mass flow rate detector 21 is taken in, and then at step 106, FIG.
The target air-fuel ratio A / F is calculated from the map shown in 2 (B). Next, at step 107, the fuel injection amount Q required to bring the air-fuel ratio to the target air-fuel ratio A / F is calculated based on the intake air amount Ga and the target air-fuel ratio A / F.

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 EGR control valve 31 immediately reach the required load L and the engine speed N. According to the target opening ST,
It is matched with SE. Therefore, for example, when the required load L is increased, the amount of air in the combustion chamber 5 is immediately increased, and thus the torque generated by the engine is immediately increased.

On the other hand, the opening of the throttle valve 20 or 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. To be done. That is, the fuel injection amount Q is changed after the intake air amount Ga actually changes.

When it is judged at 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, at step 110, the target fuel injection amount Q is calculated from the map shown in FIG. 16, and the fuel injection amount is made this target fuel injection amount Q. Then step 11
1, the throttle valve 20 is changed from the map shown in FIG.
The target opening degree ST of is calculated. Next, at step 112, the target opening degree SE of the EGR control valve 31 is calculated from the map shown in FIG. 15 (B), and the opening degree of the EGR control valve 31 is made this target opening degree SE.

Next, at step 113, the intake air amount Ga detected by the mass flow rate detector 21 is taken in. Next, at step 114, the fuel injection amount Q and the intake air amount Ga
The actual air-fuel ratio (A / F) _R is calculated from this. Next, at step 115, the target air-fuel ratio A / F is calculated from the map shown in FIG. 14 (B). Next, at step 116, it is judged if the actual air-fuel ratio (A / F) _R is larger than the target air-fuel ratio A / F. When (A / F) _R > A / F, the routine proceeds to step 117, where the correction value ΔS of the throttle opening is
T is decreased by a constant value α, then step 11
Proceed 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 to this 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 as described above, the fuel injection amount immediately matches the target fuel injection amount Q corresponding to the required load L and the engine speed N. Be punished. Request load L
Is immediately increased, the fuel injection amount is immediately increased, and thus the torque generated by 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 degree 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 routine proceeds from step 100 to step 108, where it is judged if the required load L has become lower than the second boundary Y (N). When L ≧ Y (N), step 110
And the second combustion is performed under a lean air-fuel ratio.

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

In the above-described embodiments, the fuel injection amount Q is open-loop controlled when the low temperature combustion is performed, and the air-fuel ratio changes the opening degree 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 EGR controlled when the second combustion is being performed. It can also be controlled by changing the opening degree of the valve 31.

As described above, according to this embodiment, the amount of EGR gas supplied to the combustion chamber 5 is greater than the amount of EGR gas at which the amount of soot generated reaches a peak in steps 103 to 107, and soot is almost generated. Since low temperature combustion is performed, soot is discharged from the internal combustion engine and N
It is possible to prevent Ox from being discharged at the same time.

Further, according to this embodiment, the EGR control valve 3
The rotary shaft 61 of the first valve body is arranged perpendicularly to the direction of the EGR gas flow and at the center of the valve body 60. Therefore, the force that the valve body on one side of the rotary shaft 61 of the valve body receives due to the negative pressure and the force that the valve body of the other side receives against the rotary shaft 61 of the valve body due to the negative pressure cancel each other out. It as a result,
Even under the condition that a negative pressure is generated in the air suction pipe 17, the EGR
The control valve 31 is not affected by the negative pressure, and therefore the opening degree of the EGR control valve 31 is maintained at a desired opening degree. Also, EG
Since the rotary shaft 61 of the valve body of the R control valve 31 is arranged perpendicularly to the direction of the EGR gas flow and in the center of the valve body 60, the width W1 of the EGR passage 29 is not increased.
The maximum flow passage width W2 of EGR gas can be made larger than the maximum flow passage width W2 'of the conventional stroke type EGR control valve shown in FIG. Further, as described above, since the EGR control valve 31 is less likely to be affected by the negative pressure, the drive speed of the EGR control valve 31 can be increased.

Further, also in the second embodiment described above,
The rotation shaft 561 of the valve body of the EGR control valve 531 is arranged perpendicularly to the direction of the EGR gas flow and at the center of the valve body 560. Therefore, with respect to the rotary shaft 561 of the valve body, the force which the valve body on one side receives by the negative pressure and the rotary shaft 5 of the valve body 5
The force received by the negative pressure on the valve element on the other side with respect to 61 is offset. As a result, the EGR control valve 531 is not affected by the negative pressure even under the condition that the negative pressure is generated in the air suction pipe 17, and therefore the opening degree of the EGR control valve 531 is maintained at the desired opening degree. . Further, the rotary shaft 561 of the valve body of the EGR control valve 531 is perpendicular to the direction of the EGR gas flow,
Moreover, since it is arranged in the center of the valve body 560, the EGR passage 2
The maximum flow passage width of EGR gas can be made larger than the maximum flow passage width W2 ′ of the conventional stroke type EGR control valve shown in FIG. 18 without increasing the width of 9. Further, as described above, since the EGR control valve 531 is hard to be affected by the negative pressure, the driving speed of the EGR control valve 531 can be increased.

[0096]

According to the invention described in claims 1 to 3, a negative pressure is generated in the engine intake passage while simultaneously preventing soot and NOx from being discharged from the internal combustion engine. Even in such a case, the opening of the EGR control valve can be maintained at a desired opening, and the maximum flow passage area of the recirculated exhaust gas can be increased without increasing the outer shape of the exhaust gas recirculation passage.

According to the invention described in claims 4 and 5, it is possible to prevent the exhaust gas recirculation rate from being set to the exhaust gas recirculation rate at which the amount of soot generated peaks.

According to the invention described in claim 6, it is possible to execute appropriate combustion according to the operating region.

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.

[Brief description of drawings]

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

FIG. 2 is a detailed view of a butterfly type EGR control valve.

FIG. 3 is a detailed view of a rotary EGR control valve.

FIG. 4 is a diagram showing amounts of smoke and NOx generated, and the like.

FIG. 5 is a diagram showing a combustion pressure.

FIG. 6 is a diagram showing a fuel molecule.

FIG. 7 is a diagram showing the relationship between the amount of smoke generated and the EGR rate.

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

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

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

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

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

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

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

FIG. 15 is a diagram showing a map of a target opening degree of a throttle valve or the like.

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

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

FIG. 18 is a partial cross-sectional side view of a conventional EGR control valve.

[Explanation of symbols]

5 ... Combustion chamber 17 ... Air suction pipe 29 ... EGR passage 31 ... EGR control valve

Continuation of front page (51) Int.Cl. ⁷ Identification symbol FI F02D 21/08 301 F02D 21/08 301B 301D 41/02 351 41/02 351 (72) Inventor Masato Goto 1 Toyota-cho, Toyota City, Aichi Prefecture Toyota Automobile Co., Ltd. (72) Inventor Hiroki Murata No. 1 Toyota-cho, Toyota City, Aichi Prefecture Toyota Automobile Co., Ltd. (72) Inventor Tsukasa Abe No. 1 Toyota-cho, Aichi Prefecture Toyota Motor Co., Ltd. (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, A) JP-A-9 -287528 (JP, A) (58) Fields investigated (Int.Cl. ⁷ , DB name) F02M 25/07 F02D 41/00-45/00 F02B 1/00-23/06

Claims (8)

(57) [Claims] 1. An engine exhaust passage and an engine intake passage are recirculated.
Recirculation that is connected to the exhaust gas passage and supplied to the combustion chamber
The amount of soot generated gradually increases as the amount of ring exhaust gas increases.
Reaches the peak and is supplied to the combustion chamber.
As the amount of gas in the combustion chamber is further increased,
The temperature of the gas and its surrounding gas is lower than the temperature of soot formation.
It is an exhaust gas recirculation amount control valve provided to control the amount of recirculated exhaust gas in the exhaust gas recirculation passage of an internal combustion engine in which soot is hardly generated. An exhaust gas recirculation amount control valve in which the flow passage area of the recirculated exhaust gas is changed, the rotation axis of the valve body is arranged perpendicularly to the flow direction of the recirculated exhaust gas, and at the center of the valve body. . 2. The exhaust gas recirculation amount control valve according to claim 1, wherein the exhaust gas recirculation amount control valve is a butterfly type. 3. The exhaust gas recirculation amount control valve according to claim 1, wherein the exhaust gas recirculation amount control valve is a rotary type. 4. The internal combustion engine, wherein the first combustion is combustion in which the soot is hardly generated, and the recirculation supplied to the combustion chamber more than the amount of the recirculation exhaust gas at which the soot generation amount reaches a peak. A switching means for selectively switching between the second combustion in which the amount of exhaust gas is small and the second combustion, or when switching from the first combustion to the second combustion or from the second combustion to the first combustion The exhaust gas recirculation amount control valve according to claim 1, wherein the exhaust gas recirculation rate is changed stepwise. 5. The exhaust gas recirculation rate when the first combustion is performed is approximately 55% or more, and the exhaust gas recirculation rate when the second combustion is performed is approximately 50%. The exhaust gas recirculation amount control valve according to claim 4, which is not more than a percentage. 6. The operating region of the engine is divided into a first operating region on the low load side and a second operating region on the high load side, and the first operating region is divided into the first operating region and the second operating region on the high load side.
5. The exhaust gas recirculation amount control valve according to claim 4, wherein the first combustion is performed in the operation region of 5 and the second combustion is performed in the second operation region. 7. The exhaust gas recirculation amount according to claim 1, wherein the internal combustion engine includes a catalyst having an oxidizing function for oxidizing unburned hydrocarbons discharged from the combustion chamber, in the engine exhaust passage. Control valve. 8. The catalyst is an oxidation catalyst, a three-way catalyst or NO.
8. The exhaust gas recirculation amount control valve according to claim 7, comprising at least one x-absorbent.