JP4591403B2 - Control device for internal combustion engine - Google Patents

Description

  The present invention relates to a control device for an internal combustion engine.

  For internal combustion engines mounted on automobiles, especially diesel engines (compression ignition internal combustion engines), the emission of fine particles (PM: Particulate Matter) represented by smoke and nitrogen oxides (NOx) into the atmosphere is further suppressed. Is required to do. In order to meet such demands, for example, DPF (Diesel Particulate Filter) that collects PM, NOx storage reduction type or selective reduction type NOx catalyst, DPNR (Diesel Particulate-NOx-Reduction system that simultaneously reduces PM and NOx) ) And other exhaust purification devices (post-treatment devices) are used.

  In DPF and DPNR, when PM accumulates, the back pressure of the engine increases. Therefore, it is necessary to periodically burn and remove the accumulated PM. This process is called PM regeneration. When performing PM regeneration, it is necessary to set DPF, DPNR, and the like to a high temperature of about 600 ° C.

  Further, SOx (sulfur oxide) contained in the exhaust gas is adsorbed on the NOx catalyst and DPNR. If such a sulfur content accumulates, the NOx catalyst and DPNR are poisoned and the purification capacity is reduced, so that the adsorbed sulfur content must be periodically removed. This process is called S poisoning regeneration. In order to perform S poisoning regeneration, the NOx catalyst and DPNR need to be set at a high temperature of, for example, about 650 to 700 ° C.

  Thus, when performing PM regeneration, S poison regeneration, etc. of an exhaust purification device, it is necessary to make an exhaust purification device high temperature. However, since the exhaust gas temperature is low in the light load region and the medium load region, it is normal that the exhaust purification device is not at a temperature necessary for regeneration. For this reason, when performing PM regeneration or S poison regeneration in a light and medium load range, exhaust gas purification is performed by circulating HC (unburned fuel) and oxygen in the exhaust passage and burning the HC with an exhaust purification device. The temperature of the device needs to be raised.

  As a method of circulating HC in the exhaust passage, a method of additionally injecting fuel from the injector after completion of combustion in the cylinder (post-injection), a fuel addition valve provided in the exhaust passage, or a fuel in exhaust gas There is a way to inject. However, in the case of post-injection, there is a problem that engine oil is easily diluted, and in the case of using a fuel addition valve, there is a problem that the structure is complicated and the cost is likely to increase.

  By the way, in the diesel engine, when the air-fuel ratio is lowered (riched) by increasing the EGR rate, it is known that the smoke increases initially but then decreases rapidly. This is considered to be because the combustion temperature is lowered to a temperature at which smoke is not generated by a large amount of exhaust gas recirculation (EGR). Thus, by performing a large amount of EGR, the diesel engine can be operated at an air-fuel ratio close to the stoichiometric air-fuel ratio or an air-fuel ratio richer than the stoichiometric air-fuel ratio without discharging smoke. Such a technique is hereinafter referred to as “rich combustion”.

  When performing rich combustion, unlike the case of performing normal lean combustion, a large amount of HC is contained in the exhaust gas of the diesel engine. Thus, as a method of circulating HC and oxygen in the exhaust passage, a method using this rich combustion is conceivable. In other words, if rich combustion is performed in some cylinders and normal lean combustion is performed in other cylinders, the exhaust gas from the cylinder that performs rich combustion contains a large amount of HC, and the exhaust from the cylinder that performs lean combustion The gas contains a large amount of oxygen. Therefore, if an operation that mixes rich combustion and lean combustion is performed, HC and oxygen can be circulated in the exhaust passage without using post injection or a fuel addition valve into the exhaust gas. is there.

  Japanese Patent Application Laid-Open No. 2001-152842 discloses a four-cylinder spark ignition internal combustion engine in which the first to third cylinders are lean air-fuel ratio lean cylinders and the fourth cylinder is a rich air-fuel ratio rich cylinder. A system is disclosed in which an external EGR passage is provided only in a cylinder or only in a rich cylinder, and the amount of external EGR returned to only the lean cylinder or only the rich cylinder is controlled so that the generated torques of the lean cylinder and the rich cylinder are equal. Has been.

JP 2001-152842 A JP 2000-54900 A JP 2001-329872 A JP 9-126044 A

  The system disclosed in the above publication is directed to a spark ignition internal combustion engine. On the other hand, for example, when a diesel engine is operated by being mixed with rich combustion and lean combustion, there are circumstances different from those of a spark ignition internal combustion engine. According to the knowledge of the present inventors, in this case, the following problems are considered to occur.

  In a diesel engine, it is necessary to perform EGR not only for a cylinder that performs rich combustion but also for a cylinder that performs normal lean combustion for the purpose of reducing NOx emissions. In a state where rich combustion and lean combustion are mixed and operated, the HC concentration in the exhaust gas periodically varies. That is, the HC concentration increases during the exhaust stroke of the rich combustion cylinder, and the HC concentration decreases during the exhaust stroke of the lean combustion cylinder. In this state, when a part of the exhaust gas (external EGR gas) is returned to the intake passage through the external EGR passage, the HC concentration in the intake air also changes as the HC concentration in the exhaust gas changes.

  When the HC concentration in the intake air fluctuates, the torque generated by the diesel engine also fluctuates accordingly, so that the torque fluctuation tends to increase. In particular, in the lean combustion cylinder and the rich combustion cylinder, changes in torque appear in opposite directions when the HC concentration in the intake air increases. That is, when the HC concentration in the intake air becomes high, the amount of fuel to be burned increases in the lean combustion cylinder rich in oxygen, so the torque increases, whereas in the rich combustion cylinder with little oxygen. Since it is easy to misfire, the torque decreases. For this reason, the torque of the entire engine varies greatly depending on whether the external EGR gas when the HC concentration becomes high enters the lean combustion cylinder or the rich combustion cylinder.

  The external EGR gas is mixed into the intake air with a delay required for passing through the external EGR passage, and the delay varies depending on the engine speed and the load. For this reason, the external EGR gas having a high HC concentration discharged from the rich combustion cylinder is randomly returned to the lean combustion cylinder and the rich combustion cylinder. As a result, the torque of the entire engine tends to fluctuate greatly at random. In addition, since it is extremely difficult to predict which cylinder the external EGR gas with high HC concentration will return to, it is virtually impossible to suppress such torque fluctuations by correcting the fuel injection amount. is there.

  In addition, if the external EGR passage is provided independently for each cylinder, it is possible to avoid the above-mentioned problem. However, since the structure of the intake / exhaust system becomes complicated and large, in practice, It is difficult to adopt such a configuration.

  The present invention has been made to solve the above-described problems, and is an internal combustion engine capable of suppressing torque fluctuation during execution of rich lean control in which a rich combustion cycle and a lean combustion cycle are mixed in the internal combustion engine. An object of the present invention is to provide a control device.

In order to achieve the above object, a first invention is a control device for an internal combustion engine,
An external EGR passage connecting the exhaust passage and the intake passage of the internal combustion engine;
An external EGR amount varying means for varying the amount of external EGR flowing back through the external EGR passage;
Internal EGR amount varying means for varying the amount of internal EGR generated inside the internal combustion engine without passing through the external EGR passage;
A lean combustion cycle that makes the air-fuel ratio leaner than the stoichiometric air-fuel ratio, and rich combustion that makes the air-fuel ratio relatively rich compared to the lean combustion cycle by making the EGR rate and fuel injection amount different from the lean combustion cycle A rich lean control means for performing rich lean control for periodically switching the cycle in the same cylinder or continuing in separate cylinders;
With
The rich lean control means limits the external EGR rate during execution of the rich lean control to a predetermined upper limit value or less, and covers the remaining EGR amount necessary for each of the lean combustion cycle and the rich combustion cycle with the internal EGR. As described above, it includes an EGR control means for controlling states of the external EGR amount variable means and the internal EGR amount variable means.

The second invention is the first invention, wherein
The predetermined upper limit value is the torque between the rich combustion cycle and the lean combustion cycle when the exhaust gas in the rich combustion cycle is recirculated by external EGR and mixed with the intake air so that the HC concentration in the intake air increases. The difference is a value determined so as to be within an allowable range.

The third invention is the first invention, wherein
The EGR control means sets the external EGR rate during execution of the rich lean control to substantially zero.

According to a fourth invention, in any one of the first to third inventions,
The internal EGR amount varying means includes a variable valve mechanism that varies the magnitude of positive or negative valve overlap between the intake valve and the exhaust valve for each cylinder, each cylinder group, or each cycle. Features.

According to a fifth invention, in any one of the first to fourth inventions,
An exhaust purification device disposed in the exhaust passage of the internal combustion engine;
The rich lean control means performs the rich lean control when regeneration of the exhaust purification device is required.

According to a sixth invention, in any one of the first to fifth inventions,
In accordance with the load of the internal combustion engine, the rich lean control means is configured so that each of the EGR rate and the fuel injection amount is set so that the rich combustion cycle torque and the lean combustion cycle torque become equal in the low load region. And a torque control means for controlling each EGR rate and fuel injection amount so that the torque of the lean combustion cycle is larger than the torque of the rich combustion cycle in the high load side region. .

According to a seventh invention, in any one of the first to sixth inventions,
The rich lean control means performs the lean combustion cycle and the rich combustion cycle in separate cylinders in a region of a high load side or a high exhaust gas temperature side according to a load or exhaust gas temperature of the internal combustion engine. And a mode switching means for switching the lean combustion cycle and the rich combustion cycle periodically in the same cylinder in the low load side or the low exhaust gas temperature side region. .

Further, an eighth invention is any one of the first to seventh inventions,
A variable valve mechanism that makes the intake valve closing timing variable;
Rich combustion cycle intake valve slow closing means for delaying the intake valve closing timing in the rich combustion cycle compared to the intake valve closing timing in the lean combustion cycle when the rich lean control is executed;
Is further provided.

  According to the first invention, in the internal combustion engine, the air-fuel ratio is made relatively rich by making the air-fuel ratio leaner than the stoichiometric air-fuel ratio and making the EGR rate and fuel injection amount different from the lean combustion cycle. It is possible to perform rich lean control in which the rich combustion cycle to be performed is periodically switched in the same cylinder or continued in different cylinders. When the rich lean control is executed, the external EGR rate is limited to a predetermined upper limit value or less, and the external EGR amount variable means and the internal EGR cover the remainder of the EGR amount necessary for each of the lean combustion cycle and the rich combustion cycle; The state of the internal EGR amount varying means can be controlled. When the rich lean control is executed, the HC concentration in the external EGR gas varies periodically. The torque increases when the external EGR gas having a high HC concentration flows into the lean combustion cycle, whereas the torque decreases when the external EGR gas having a high HC concentration flows into the rich combustion cycle. For this reason, when the external EGR rate at the time of executing the rich lean control is high, there is a problem that the torque fluctuation tends to increase. According to the first aspect of the present invention, it is possible to reliably prevent the torque fluctuation of the internal combustion engine from becoming excessive by limiting the external EGR rate at the time of executing the rich lean control.

  According to the second invention, when the exhaust gas in the rich combustion cycle is recirculated by the external EGR and mixed with the intake air when the rich HC control is executed, the upper limit value of the external EGR rate at the time of executing the rich lean control is increased. In addition, the torque difference between the rich combustion cycle and the lean combustion cycle can be determined to be within an allowable range. Thereby, the torque fluctuation of the internal combustion engine when the rich lean control is executed can be surely suppressed to a value below the allowable value.

  According to the third aspect of the invention, the external EGR rate during execution of rich lean control can be made substantially zero. Thereby, the torque fluctuation of the internal combustion engine caused by the external EGR at the time of executing the rich lean control can be completely eliminated, so that the torque fluctuation can be particularly reduced.

  According to the fourth invention, the internal EGR amount varying means is a variable valve that varies the magnitude of the positive or negative valve overlap between the intake valve and the exhaust valve for each cylinder, for each cylinder group, or for each cycle. It can be configured with a mechanism. As a result, the internal EGR amount can be accurately controlled in each of the rich combustion cycle and the lean combustion cycle.

  According to the fifth aspect, rich lean control can be executed when regeneration of the exhaust purification device disposed in the exhaust passage of the internal combustion engine is required. According to the rich lean control, it is possible to regenerate the exhaust purification device in a wide operating range without taking post injection or adding a fuel to the exhaust gas.

  According to the sixth aspect of the present invention, when the rich lean control is executed, the EGR rate and the fuel injection amount are controlled so that the rich combustion cycle torque and the lean combustion cycle torque become equal in the low load side region. In the region on the high load side, each EGR rate and fuel injection amount can be controlled so that the torque of the lean combustion cycle becomes larger than the torque of the rich combustion cycle. As a result, the torque fluctuation of the internal combustion engine can be further reduced in the low load side region, and rich combustion can be performed in the high load side region.

  According to the seventh aspect, in the high load side or high exhaust gas temperature side region, the lean combustion cycle and the rich combustion cycle are continued in separate cylinders, respectively, and the low load side or low exhaust gas temperature side region In, the lean combustion cycle and the rich combustion cycle can be switched periodically in the same cylinder. As a result, it is possible to reliably prevent the in-cylinder temperature of the rich combustion cycle from becoming too high in the region on the high load side or the high exhaust gas temperature side, and it is possible to reliably prevent smoke discharge. In addition, in the region of the low load side or the low exhaust gas temperature side, it is possible to reliably avoid the in-cylinder temperature of the rich combustion cycle becoming too low, and to suppress the occurrence of misfire and stabilize the combustion. Can do.

  According to the eighth aspect of the invention, when the rich lean control is executed, the intake valve closing timing in the rich combustion cycle can be delayed as compared with the intake valve closing timing in the lean combustion cycle. As a result, it is possible to reliably avoid the in-cylinder temperature of the rich combustion cycle from becoming too high, and to reliably prevent smoke from being discharged.

  Embodiments of the present invention will be described below with reference to the drawings. In addition, the same code | symbol is attached | subjected to the element which is common in each figure, and the overlapping description is abbreviate | omitted.

Embodiment 1 FIG.
[Description of system configuration]
FIG. 1 is a diagram for explaining a system configuration according to the first embodiment of the present invention. The system shown in FIG. 1 includes a four-cycle diesel engine (compression ignition internal combustion engine) 10. It is assumed that the diesel engine 10 is mounted on a vehicle and used as a power source. The diesel engine 10 of the present embodiment is an in-line four-cylinder type having four cylinders # 1 to # 4, but the number of cylinders and the cylinder arrangement of the diesel engine in the present invention are not limited to this. Absent.

  Each cylinder of the diesel engine 10 is provided with an injector 12 that injects fuel directly into the cylinder. The injectors 12 of each cylinder are connected to a common common rail 14. Fuel in a fuel tank (not shown) is pressurized to a predetermined fuel pressure by a supply pump 16, stored in the common rail 14, and supplied from the common rail 14 to each injector 12.

  An exhaust passage 18 of the diesel engine 10 is branched by an exhaust manifold 20 and connected to an exhaust port 22 (see FIG. 2) of each cylinder. The diesel engine 10 according to this embodiment includes a turbocharger 24. The exhaust passage 18 is connected to the exhaust turbine of the turbocharger 24.

  An exhaust purification device 26 that purifies exhaust gas is provided in the exhaust passage 18 downstream of the turbocharger 24. As the exhaust purification device 26, for example, one of an oxidation catalyst, a NOx storage reduction type or selective reduction type NOx catalyst, a DPF (Diesel Particulate Filter), a DPNR (Diesel Particulate-NOx-Reduction system), or a combination thereof Etc. can be used.

  An air cleaner 30 is provided near the inlet of the intake passage 28 of the diesel engine 10. The air drawn through the air cleaner 30 is compressed by the intake compressor of the turbocharger 24 and then cooled by the intercooler 32. The intake air that has passed through the intercooler 32 is distributed by the intake manifold 34 to the intake ports 35 (see FIG. 2) of the respective cylinders.

  An intake throttle valve 36 is installed between the intercooler 32 and the intake manifold 34 in the intake passage 28. An air flow meter 38 for detecting the amount of intake air is installed in the vicinity of the intake passage 28 downstream of the air cleaner 30.

  One end of an external EGR passage 40 is connected to the intake passage 28 in the vicinity of the intake manifold 34. The other end of the external EGR passage 40 is connected to the vicinity of the exhaust manifold 20 of the exhaust passage 18. In this system, a part of the exhaust gas (burned gas) can be recirculated to the intake passage 28 through the external EGR passage 40, that is, external EGR (Exhaust Gas Recirculation) can be performed. Hereinafter, the exhaust gas recirculated through the external EGR passage 40 is referred to as “external EGR gas”.

  In the middle of the external EGR passage 40, an EGR cooler 42 for cooling the external EGR gas is provided. An EGR valve 44 is provided downstream of the EGR cooler 42 in the external EGR passage 40. By changing the opening degree of the EGR valve 44, the amount of exhaust gas passing through the external EGR passage 40, that is, the amount of external EGR can be adjusted.

  An intake pressure sensor 46 that detects the intake pressure is installed downstream of the intake throttle valve 36 in the intake passage 28. An exhaust temperature sensor 47 that detects the temperature of the exhaust gas is installed in the exhaust passage 18. Furthermore, this system includes an accelerator opening sensor 48 that detects the amount of depression of the accelerator pedal (accelerator opening).

  The system of this embodiment includes an ECU (Electronic Control Unit) 50. The ECU 50 is connected to the various sensors and actuators described above. The ECU 50 controls the operating state of the diesel engine 10 by driving each actuator according to a predetermined program based on the output of each sensor.

  FIG. 2 is a view showing a cross section of one cylinder of the diesel engine 10 in the system shown in FIG. Hereinafter, the diesel engine 10 will be further described. As shown in FIG. 2, a crank angle sensor 62 that detects the rotation angle (crank angle) of the crankshaft 60 is attached in the vicinity of the crankshaft 60 of the diesel engine 10. The crank angle sensor 62 is connected to the ECU 50.

  Further, the diesel engine 10 is provided with an intake variable valve mechanism 54 that drives the intake valve 52 and an exhaust variable valve mechanism 58 that drives the exhaust valve 56. These intake variable valve mechanism 54 and exhaust variable valve mechanism 58 are connected to the ECU 50. Specific configurations of the intake variable valve mechanism 54 and the exhaust variable valve mechanism 58 are not particularly limited. For example, an electromagnetically driven valve or a hydraulically driven valve that can be opened and closed at an arbitrary timing can be used. Alternatively, a mechanical mechanism such as a mechanism that continuously varies the phase of a cam (not shown) that drives the intake valve 52 and the exhaust valve 56 may be used.

[Features of Embodiment 1]
(Internal EGR with valve overlap)
In the diesel engine 10 of this embodiment, the magnitude of the negative valve overlap between the intake valve 52 and the exhaust valve 56 can be continuously changed by the intake variable valve mechanism 54 and the exhaust variable valve mechanism 58. . FIG. 3 is a diagram for explaining negative valve overlap. As shown in FIG. 3, the negative valve overlap is a state where both the intake valve 52 and the exhaust valve 56 are closed after the exhaust valve 56 is closed until the intake valve 52 is opened.

  3 is a valve lift diagram of the intake valve 52 and the exhaust valve 56 when no negative valve overlap is provided. From this state, the closing timing of the exhaust valve 56 is advanced and the opening timing of the intake valve 52 is delayed, so that a negative valve overlap can be caused. A thick curve in FIG. 3 is a valve lift diagram of the intake valve 52 and the exhaust valve 56 when a negative valve overlap is generated. The magnitude (period) of the negative valve overlap can be changed according to the degree to which the closing timing of the exhaust valve 56 and the opening timing of the intake valve 52 are changed.

  When a negative valve overlap is generated, the exhaust valve 56 is closed before the burned gas in the cylinder completely flows out to the exhaust port 22. The burned gas that has not been discharged to the exhaust port 22 remains in the cylinder as it is, or once exits to the intake port 35 when the intake valve 52 is opened, and then the piston 64 descends together with fresh air. Inhaled. When a negative valve overlap occurs, internal EGR can be performed in this way. As the negative valve overlap is increased, the internal EGR amount can be increased. Hereinafter, the exhaust gas recirculated by the internal EGR is referred to as “internal EGR gas”.

  As shown in FIG. 3, in the present embodiment, the opening timing of the exhaust valve 56 and the closing timing of the intake valve 52 are also changed as the magnitude of the negative valve overlap is changed. The magnitude of the negative valve overlap may be changed without changing the opening timing of the exhaust valve 56 and the closing timing of the intake valve 52. In this embodiment, the magnitude of the negative valve overlap is changed by changing both the closing timing of the exhaust valve 56 and the opening timing of the intake valve 52. However, the closing timing of the exhaust valve 56 is changed. The magnitude of the negative valve overlap may be changed by changing either the opening timing of the intake valve 52 or the intake valve 52. In this case, the diesel engine 10 may include only one of the intake variable valve mechanism 54 and the exhaust variable valve mechanism 58. Furthermore, in the present invention, internal EGR is performed by generating a normal valve overlap (positive valve overlap) in which both the intake valve 52 and the exhaust valve 56 are opened, and the magnitude of the positive valve overlap. The internal EGR amount may be adjusted by changing the length.

  As described above, the diesel engine 10 of the present embodiment can perform external EGR and internal EGR. By performing EGR, the combustion temperature can be lowered, so that NOx emission can be suppressed.

(Rich combustion)
In such a diesel engine 10, the ratio of the EGR gas, which is the sum of the external EGR gas and the internal EGR gas, out of the gas flowing into the cylinder, that is, when the EGR rate is increased, the air-fuel ratio A / F decreases (enriches). In this case, the air-fuel ratio A / F is the sum of the amount of fresh air that has flowed into the cylinder and the amount of air remaining in the EGR gas that has flowed into the cylinder (the amount of air corresponding to the unconsumed oxygen amount). It is determined by the total air amount and the fuel injection amount from the injector 12.

  The diesel engine 10 can perform combustion with the in-cylinder temperature lowered to a temperature at which smoke is not generated by increasing the EGR rate and performing a large amount of EGR. Such combustion is referred to herein as “rich combustion”. The air-fuel ratio A / F in rich combustion is significantly reduced (riched) compared to normal combustion (lean combustion) of a diesel engine.

  In the present invention, the air-fuel ratio A / F at the time of rich combustion may be a value below the stoichiometric air-fuel ratio, that is, the stoichiometric air-fuel ratio or a richer value, but a value larger than the stoichiometric air-fuel ratio, that is, the theoretical air-fuel ratio. It may be a value leaner than the air-fuel ratio (for example, about 15 to 17). In other words, “rich” in “rich combustion” does not mean “richer than the stoichiometric air-fuel ratio”, but means “relatively rich” compared to lean combustion described later.

(Rich lean control)
If the EGR rate is made lower than the rich combustion and the amount of air in the cylinder is increased, the air-fuel ratio A / F becomes larger (lean). In the present specification, combustion in which the air-fuel ratio A / F is larger than the air-fuel ratio A / F of rich combustion and larger than the stoichiometric air-fuel ratio is referred to as “lean combustion”. In the system of this embodiment, the diesel engine 10 can be operated by mixing a rich combustion cycle that performs rich combustion and a lean combustion cycle that performs lean combustion. Hereinafter, the control for mixing the rich combustion cycle and the lean combustion cycle is referred to as “rich lean control”.

  FIG. 4 is a diagram showing the relationship between the air-fuel ratio A / F and the concentration of HC contained in the exhaust gas discharged by combustion at the air-fuel ratio A / F. As shown in the figure, the lower the air-fuel ratio A / F, the higher the HC concentration in the exhaust gas. For this reason, a large amount of HC is contained in the exhaust gas of rich combustion. On the other hand, the lean combustion exhaust gas is rich in oxygen. For this reason, exhaust gas containing abundant HC and oxygen can be circulated in the exhaust passage 18 by performing rich lean control.

  FIG. 5 is a diagram for explaining an example of a combustion mode of each cylinder when rich lean control is performed. In the present embodiment, when rich lean control is performed, the rich combustion cycle and the lean combustion cycle may be switched periodically in the same cylinder, or may be continued in separate cylinders. .

  Pattern 1 in FIG. 5 is an example in which the rich combustion cycle and the lean combustion cycle are continued in separate cylinders, and among the four cylinders of the diesel engine 10, the rich combustion cycle is in the # 1 and # 4 cylinders. Are continuously performed, and the lean combustion cycle is continuously performed in the # 2 and # 3 cylinders.

  In this case, the number of cylinders that continue the rich combustion cycle may not be the same as the number of cylinders that continue the lean combustion cycle. For example, as shown in pattern 2 in FIG. 5, the rich combustion cycle may be continuously performed in # 1, and the lean combustion cycle may be continuously performed in # 2, # 3, and # 4 cylinders.

  In these cases, by increasing the internal EGR amount of the cylinder that continues the rich combustion cycle more than the internal EGR amount of the cylinder that continues the lean combustion cycle, the EGR rate of the former cylinder is greater than the EGR rate of the latter cylinder. You should make it high. Therefore, in this case, the intake variable valve mechanism 54 and the exhaust variable valve mechanism 58 are negative valves for the cylinder or the cylinder group that continues the rich combustion cycle and the cylinder or the cylinder group that continues the lean combustion cycle. It is sufficient that the size of the overlap can be changed separately.

  On the other hand, pattern 3 in FIG. 5 is an example in which the rich combustion cycle and the lean combustion cycle are periodically switched in the same cylinder. In the # 1 and # 4 cylinders, the rich combustion cycle and the lean combustion cycle are changed. It is performed alternately (that is, by switching every cycle). In the # 2 and # 3 cylinders, the lean combustion cycle is continuously performed.

  In this case, switching between the rich combustion cycle and the lean combustion cycle may be performed every plural cycles instead of every cycle. Further, as shown in pattern 4 in FIG. 5, the rich combustion cycle and the lean combustion cycle may be periodically switched in all the cylinders.

  In the case where the rich combustion cycle and the lean combustion cycle are switched for each cycle in the same cylinder, the EGR rate may be changed for each cycle by changing the internal EGR amount for each cycle. That is, in this case, the intake variable valve mechanism 54 and the exhaust variable valve mechanism 58 change the magnitude of the negative valve overlap of the cylinder that switches between the rich combustion cycle and the lean combustion cycle for each cycle. It only has to come to be able to.

  By the way, when performing rich lean control, in order to make the torque fluctuation of the diesel engine 10 as small as possible, it is preferable to make the rich combustion cycle torque equal to the lean combustion cycle torque. On the other hand, since a large amount of EGR is required to establish rich combustion, the amount of air that can be sucked into the cylinder is reduced accordingly. Therefore, in the rich combustion cycle, the upper limit of the amount of fuel that can be combusted in the cylinder is lower than in the lean combustion cycle (normal cycle), so the upper limit of the torque that can be generated is also lower.

  Therefore, in the present embodiment, when rich lean control is performed, if the torque required for the diesel engine 10 is equal to or less than the value obtained by multiplying the upper limit torque of the rich combustion cycle by the number of all cylinders, the torque of the rich combustion cycle And the EGR rate and the fuel injection amount are controlled so that the torque of the lean combustion cycle becomes equal.

  In order to make the torque of the rich combustion cycle equal to the torque of the lean combustion cycle, it is necessary to make the fuel injection amount of the rich combustion cycle larger than the fuel injection amount of the lean combustion cycle. This is because in the rich combustion cycle, the combustion becomes slow and the thermal efficiency is lowered.

  On the other hand, when the rich lean control is performed, if the torque required for the diesel engine 10 exceeds the value obtained by multiplying the upper limit torque of the rich combustion cycle by the number of all cylinders, the upper limit torque is generated in the rich combustion cycle. At the same time, in the lean combustion cycle, each EGR rate and fuel injection amount are controlled so as to generate torque exceeding the upper limit torque of the rich combustion cycle. By doing in this way, even if the torque required for the diesel engine 10 exceeds the value obtained by multiplying the upper limit torque of the rich combustion cycle by the number of all cylinders, that is, the middle and high load regions, rich combustion is performed. It can be carried out.

  The rich lean control as described above has the following advantages. As a first advantage, the temperature of the exhaust purification device 26 can be raised in a light and medium load range. When performing PM regeneration that burns and removes PM (Particulate Matter) accumulated in the exhaust purification device 26, or when performing S poisoning regeneration that desorbs the sulfur adsorbed by the exhaust purification device 26, exhaust gas Although it is necessary to make the purification device 26 in a high temperature state, the exhaust purification device 26 is usually not in such a high temperature state in a light and medium load region where the exhaust temperature is low. In the present invention, by performing rich lean control in such a case, a large amount of HC and oxygen is supplied to the exhaust purification device 26, and the HC is burned by the exhaust purification device 26. The temperature can be raised to a temperature at which PM regeneration or S poisoning regeneration can be performed.

  As described above, when rich lean control is performed for the purpose of raising the temperature of the exhaust purification device 26, the air-fuel ratio A / F of the rich combustion cycle may be richer than the stoichiometric air-fuel ratio, or leaner than the stoichiometric air-fuel ratio. Good.

  Note that if the air-fuel ratio A / F of the rich combustion cycle is leaner than the stoichiometric air-fuel ratio, oxygen is also contained in the exhaust gas of the rich combustion cycle itself, so the rich combustion cycle continues for all cylinders. Even so, oxygen and a large amount of HC are supplied to the exhaust purification device 26. For this reason, even in this case, it seems that the exhaust gas purification device 26 can be raised in temperature if simply considered. However, in this case, since the oxygen concentration as a whole is low, HC cannot be efficiently burned by the exhaust purification device 26, and as a result, the temperature of the exhaust purification device 26 cannot be sufficiently increased. On the other hand, when rich lean control is performed, since a large amount of oxygen can be supplied to the exhaust purification device 26 by mixing lean combustion cycles, HC can be burned with high efficiency, The temperature of the purification device 26 can be raised sufficiently.

  As a second advantage of the rich lean control, S poisoning regeneration in a medium and high load range is possible. When performing S poisoning regeneration, it is required to supply rich exhaust gas having a stoichiometric air-fuel ratio or less to the exhaust purification device 26. However, if it is assumed that the rich combustion cycle is continued in all cylinders, the above-described operation is performed. As described above, the torque that can be generated by the diesel engine 10 is limited to a value obtained by multiplying the upper limit torque of the rich combustion cycle by the number of all cylinders. For this reason, in the middle and high load range, it is often impossible to perform the S poison regeneration by the continuous operation of the rich combustion in all cylinders. On the other hand, according to the rich lean control, rich exhaust gas can be supplied by the rich combustion cycle while gaining a large torque in the lean combustion cycle, so that it is possible to regenerate S poisoning in the middle and high load range. . Note that, during S poisoning regeneration, the air-fuel ratio A / F of the rich combustion cycle is made richer than the theoretical air-fuel ratio.

  As described above, according to the rich lean control, it is possible to expand the operation range in which the exhaust gas purification device 26 can perform the PM regeneration and the S poison regeneration. Furthermore, there is no adverse effect as in the case of performing PM regeneration or S poison regeneration by a method such as post injection or fuel addition to exhaust gas. That is, there is an advantage that problems such as dilution of engine oil by post-injection and cost increase due to installation of a fuel addition valve in the exhaust passage do not occur.

(Phenomenon caused by external EGR during rich lean control)
If the external EGR is performed in parallel with the internal EGR during the execution of the rich lean control as described above, the following problems occur. As shown in FIG. 4, the HC concentration in the exhaust gas of the rich combustion cycle is high, and the HC concentration in the exhaust gas of the lean combustion cycle is low. For this reason, during rich lean control, the HC concentration in the exhaust gas periodically varies. That is, the HC concentration is high during the exhaust stroke of the rich combustion cycle, and the HC concentration is low during the exhaust stroke of the lean combustion cycle. In this state, when a part of the exhaust gas (external EGR gas) is returned to the intake passage 28 through the external EGR passage, the HC concentration in the intake air also varies as the HC concentration in the exhaust gas varies.

  FIG. 6 is a diagram illustrating a change in generated torque when the fuel injection amount from the injector 12 is constant and the HC concentration in the intake air is changed. As shown in the figure, when the HC concentration in the intake air becomes high, the torque changes in the opposite direction between the lean combustion cycle with a large air-fuel ratio A / F and the rich combustion cycle with a small air-fuel ratio A / F. Appears in That is, in the lean combustion cycle in which the cylinder is rich in oxygen, the higher the HC concentration in the intake air, the more fuel is burned, and the torque increases. In contrast, in a rich combustion cycle in which the in-cylinder oxygen is scarce, the higher the HC concentration in the intake air, the easier it is to misfire, and the torque decreases. For this reason, the torque of the diesel engine 10 varies greatly depending on whether the external EGR gas when the HC concentration becomes high enters the cylinder that performs the lean combustion cycle or the cylinder that performs the rich combustion cycle. Become.

  The external EGR gas is mixed into the intake air with a delay required for passing through the external EGR passage 40, and the delay varies depending on the engine speed and the load. For this reason, the external EGR gas having a high HC concentration discharged from the cylinder that has performed the rich combustion cycle is randomly returned to the cylinder that performs the lean combustion cycle and the cylinder that performs the rich combustion cycle. As a result, there arises a problem that the torque of the diesel engine 10 tends to fluctuate greatly at random.

  In the case of the external EGR, it is extremely difficult to predict which cylinder the external EGR gas having a high HC concentration will return to. Therefore, the random torque fluctuation as described above is corrected by correcting the fuel injection amount, etc. It is virtually impossible to suppress by.

  Therefore, in the present embodiment, in order to solve the problem of torque fluctuation as described above, when the rich lean control is executed, the external EGR is stopped, and all the EGR amount (EGR rate) required for each cylinder is The internal EGR for each cylinder was covered.

  In the case of internal EGR, the internal EGR gas flowing into the cylinder is the exhaust gas of the previous cycle performed in that cylinder. That is, in the case of the internal EGR, there is no problem of time delay like the external EGR and the problem of unpredictable which cylinder flows into. For this reason, in the case of internal EGR, it is possible to predict the HC concentration in the internal EGR gas flowing into the cylinder.

  That is, if the EGR at the time of rich lean control is covered by the internal EGR, the HC concentration in the intake air from the EGR rate and the HC concentration in the internal EGR gas determined according to the air-fuel ratio A / F of the cylinder. Can be determined. If the HC concentration in the intake air is determined, it is possible to set the fuel injection amount so that the torque of the cycle matches the target torque. Therefore, according to the present embodiment, the torque of each cylinder (each cycle) can be accurately matched with the target torque when the rich lean control is executed, and the occurrence of large torque fluctuations can be avoided. It becomes possible.

[Specific Processing in Embodiment 1]
FIG. 7 is a flowchart of a routine executed by the ECU 50 in the present embodiment in order to realize the above function. This routine is repeatedly executed every predetermined time.

  According to the routine shown in FIG. 7, it is first determined whether or not there is a request for rich lean control (step 100). In the present embodiment, a rich lean control request is issued when PM regeneration or S poisoning regeneration is performed in a light and medium load region, or when S poisoning regeneration is performed in a medium or high load region. If it is determined in step 100 that a rich lean control request has not been issued, the current processing cycle is terminated.

  On the other hand, if it is determined that the rich lean control request has been issued, then the operating state of the diesel engine 10 such as the engine speed, the accelerator opening, the intake air amount, the intake pressure, etc. Detection is performed based on the sensor signal (step 102).

  Next, based on the operating state detected in step 102, the magnitude of the negative valve overlap, the fuel injection amount, and the target value of the fuel injection timing in the rich combustion cycle during the rich lean control are acquired (step). 104). The ECU 50 is a map that defines the relationship between the target value of the negative valve overlap size, fuel injection amount, and fuel injection timing in the rich combustion cycle, and the operating state of the diesel engine 10 (hereinafter, “rich combustion map”). Is stored). In step 104, the target value of the rich combustion cycle negative valve overlap, fuel injection amount, and fuel injection timing corresponding to the current operating state is acquired by referring to the rich combustion map. The

  Subsequently, based on the operating state detected in step 102, the negative valve overlap magnitude, fuel injection amount, and fuel injection timing target values in the lean combustion cycle during rich lean control are acquired ( Step 106). The ECU 50 is a map that defines the relationship between the target value of the negative valve overlap size, fuel injection amount, and fuel injection timing in the lean combustion cycle and the operating state of the diesel engine 10 (hereinafter referred to as “lean combustion map”). Is stored). In step 106, by referring to the lean combustion map, the magnitude of the negative valve overlap, the fuel injection amount, and the fuel injection timing target value corresponding to the current operating state are acquired. The

The rich combustion map and the lean combustion map are set so as to follow the following rules (1) to (3), respectively.
(1) The magnitude of the negative valve overlap is such that the EGR rate necessary to achieve the target air-fuel ratio can be obtained according to the target air-fuel ratio of each of the rich combustion cycle and the lean combustion cycle. Is set to That is, the negative valve overlap of the rich combustion cycle is set to be larger than the negative valve overlap of the lean combustion cycle.
(2) The fuel injection amount is the HC concentration in the intake air determined from the EGR rate and the HC concentration in the internal EGR gas determined according to the air-fuel ratio A / F of the rich combustion cycle and the lean combustion cycle. The injection amount is set so that a target torque for each cycle determined as described later can be obtained in consideration of the influence of the torque on the torque. That is, in consideration of the relationship as shown in FIG. 6, the fuel injection amount is set to be smaller so that the influence of the HC in the intake is canceled as the HC concentration in the intake becomes higher.
(3) For the target torque, when the required torque to the diesel engine 10 is an operating region on the low load side such that the upper limit torque of the rich combustion cycle is equal to or less than the value of all the cylinders, The target torque for the rich combustion cycle and the target torque for the lean combustion cycle are set to the same value. On the other hand, when the required torque to the diesel engine 10 is in the high load side operating region such that the upper limit torque of the rich combustion cycle is multiplied by the total number of cylinders, the target torque of the rich combustion cycle is The target torque for the lean combustion cycle is set to a torque that exceeds the upper limit torque for the rich combustion cycle, and is set such that the average torque of the diesel engine 10 matches the required torque.

  More specifically, the rules (1) and (2) are set in consideration of the following circumstances.

  In a cylinder that continuously performs a rich combustion cycle or a cylinder that continuously performs a lean combustion cycle, the air-fuel ratio A / F for each cycle is constant unless the operating state changes, so the HC concentration in the internal EGR gas Is also constant. Further, since the EGR rate is also constant, the HC concentration in the intake air is also constant.

  On the other hand, in the cylinder in which the rich combustion cycle and the lean combustion cycle are alternately performed, the air-fuel ratio A / F and the EGR rate regularly vary from cycle to cycle, so the HC concentration in the internal EGR gas and the HC in the intake air Concentration also varies regularly from cycle to cycle. In the rich combustion cycle, the exhaust gas of the lean combustion cycle performed immediately before is recirculated as internal EGR gas. Therefore, a large amount of unconsumed air (oxygen) is contained in the internal EGR gas recirculated to the rich combustion cycle. Therefore, in the rich combustion cycle after the lean combustion cycle, less air (fresh air) flows into the cylinder in order to achieve the same target air-fuel ratio compared to the case where the rich combustion cycle is continuously performed. There is a need to. That is, it is necessary to further increase the EGR rate by increasing the negative valve overlap.

  When the target values of the negative valve overlap magnitude, the fuel injection amount, and the fuel injection timing are acquired for each of the rich combustion cycle and the lean combustion cycle by the processing of steps 104 and 106 described above, Rich lean control is executed using the target value (step 108). That is, the operation of the injector 12, the intake variable valve mechanism 54 and the exhaust variable valve mechanism 58 of each cylinder is controlled to realize the predetermined operation mode as exemplified in FIG. The diesel engine 10 is operated with a lean combustion cycle. In this step 108, the EGR valve 44 is fully closed and the external EGR is stopped with the execution of the rich lean control.

  According to the processing of the routine shown in FIG. 7 described above, the external EGR rate is stopped when the rich lean control is executed, and all EGR required in the rich combustion cycle and the lean combustion cycle is covered by the internal EGR. it can. For this reason, there is no possibility that a large torque fluctuation occurs due to the external EGR gas whose HC concentration fluctuates periodically flowing into each cylinder at random. Therefore, the torque fluctuation of the diesel engine 10 at the time of rich lean control execution can be effectively suppressed.

  In the first embodiment described above, the external EGR is stopped when the rich lean control is executed, that is, the external EGR rate is set to zero. However, in the present invention, the torque fluctuation of the diesel engine 10 is less than the allowable value. If the external EGR rate is within the range, the external EGR may be performed when the rich lean control is executed. The external EGR rate that can be allowed when the rich lean control is executed can be determined as follows.

First, the allowable value of the torque difference between the rich combustion cycle and the lean combustion cycle is calculated from the allowable value of torque fluctuation of the diesel engine 10. Here, it is assumed that this allowable torque difference is as large as shown in FIG. 6, and the value of the HC concentration in the intake air when the torque difference between both cycles matches the allowable torque difference is a. In this case, if the HC concentration in the intake air always falls below a, the torque fluctuation of the diesel engine 10 always falls below the allowable value. The HC concentration in the intake air becomes the highest when the exhaust gas in the rich combustion cycle is mixed with the intake air as the external EGR gas. Therefore, when the value of the HC concentration in the exhaust gas of the rich combustion cycle is b (see FIG. 4) and the allowable external EGR rate value at the time of rich lean control execution is c [%], the allowable external EGR rate value c Can be represented by the following formula.
c = a / b × 100

  As described above, according to the present invention, it is not necessary to completely stop the external EGR when the rich lean control is executed, and the EGR valve 44 is opened so that the external EGR rate when the rich lean control is executed is equal to or less than the allowable external EGR rate. The degree of external EGR may be limited by controlling the degree.

  In the first embodiment described above, the EGR valve 44 is the “external EGR amount varying means” in the first invention, and the intake variable valve mechanism 54 and the exhaust variable valve mechanism 58 are the first invention. The “allowable external EGR rate” corresponds to the “predetermined upper limit value” in the second invention, and “internal EGR amount variable means” and the “variable valve mechanism” in the fourth invention. Further, when the ECU 50 executes the processing of the routine shown in FIG. 7, the “rich lean control means” in the first invention executes the processing of step 108, thereby causing the first and third inventions to The “torque control means” according to the sixth aspect of the present invention is realized by the “EGR control means” executing the processing of steps 104 and 106 described above.

Embodiment 2. FIG.
Next, the second embodiment of the present invention will be described with reference to FIG. 8 and the like. However, the difference from the first embodiment will be mainly described, and the description of the same matters will be omitted or omitted. Simplify. The present embodiment can be realized by causing the ECU 50 to execute a routine shown in FIG. 8 described later, instead of the routine shown in FIG. 7 described above, using the hardware configuration shown in FIGS. 1 and 2. .

[Features of Embodiment 2]
In rich combustion, as described above, the in-cylinder temperature is greatly reduced by a large amount of EGR, thereby suppressing the generation of smoke. Therefore, in order to prevent smoke from being discharged in rich combustion, it is important to prevent the in-cylinder temperature from rising to a temperature at which smoke is generated.

  When EGR is performed, the combustion temperature usually decreases. However, in a high load region where the exhaust gas temperature is high, the higher the EGR rate (internal EGR rate), the higher the proportion of hot EGR gas in the cylinder. The internal temperature tends to be high. For this reason, in the rich combustion, if the EGR rate becomes too high, the in-cylinder temperature may become too high.

  In rich combustion, it is also undesirable that the in-cylinder temperature becomes too low. This is because if the in-cylinder temperature becomes too low, the combustion becomes unstable and misfires easily occur. In the low load region where the fuel injection amount is small, the in-cylinder temperature of the rich combustion may become too low.

  As described in the first embodiment, in the rich combustion cycle of the cylinder in which the rich combustion cycle and the lean combustion cycle are alternately performed, the exhaust gas in the lean combustion cycle recirculates as the internal EGR gas. Consumed air (oxygen) in large quantities. For this reason, in the rich combustion cycle of the cylinder in which the rich combustion cycle and the lean combustion cycle are alternately performed, the EGR rate is increased even if the target air-fuel ratio is the same as in the case where the rich combustion cycle is continuously performed in the same cylinder. It is necessary to further increase the air to flow into the cylinder.

  In other words, when the rich combustion cycle is continuously performed in the same cylinder, the EGR rate of the rich combustion cycle is made lower than when the rich combustion cycle and the lean combustion cycle are alternately performed in the same cylinder. Can do.

  Therefore, in the present embodiment, in the region on the high load side where the rich combustion in-cylinder temperature may be too high, the rich combustion cycle is continuously performed in the same cylinder when the rich lean control is executed. That is, the diesel engine 10 is operated in the operation mode as exemplified by the pattern 1 and the pattern 2 in FIG. Thereby, since the EGR rate of rich combustion can be made relatively low, it is possible to surely avoid that the in-cylinder temperature at the time of rich combustion becomes too high, and it is possible to reliably prevent smoke discharge. .

  On the other hand, in the present embodiment, in the low load region where the in-cylinder temperature of rich combustion may be too low, the rich combustion cycle and the lean combustion cycle are alternately switched in the same cylinder when the rich lean control is executed. I decided to do it. That is, the diesel engine 10 is operated in the operation mode exemplified by the pattern 3 and the pattern 4 in FIG. As a result, the EGR rate of the rich combustion can be made relatively high, so that the in-cylinder temperature during the rich combustion can be reliably avoided, and the combustion is stabilized by suppressing the occurrence of misfire. Can be made.

[Specific Processing in Second Embodiment]
FIG. 8 is a flowchart of a routine executed by the ECU 50 in the present embodiment in order to realize the above function. In FIG. 8, the same steps as those shown in FIG. 7 are denoted by the same reference numerals, and the description thereof is omitted or simplified.

  According to the routine shown in FIG. 8, first, as in the first embodiment, it is determined whether there is a request for rich lean control (step 100), and it is determined that a rich lean control request has been issued. If so, the operating state of the diesel engine 10 is detected (step 102).

  Subsequently, it is determined whether the region is on the high load side or on the low load side based on the operating state detected in step 102 (step 110). Here, the cylinder temperature of the rich combustion cycle is not likely to become too high or too low, and when the load exceeds that with a certain load value, it is determined as the high load side region, When the load is less than that, it is determined that the region is on the low load side.

  When it is determined in step 110 that the region is on the high load side, the rich combustion cycle and the lean combustion cycle are continued in separate cylinders, that is, as exemplified by pattern 1 and pattern 2 in FIG. A mode is selected (step 112). On the other hand, when it is determined in step 110 that the region is on the low load side, a mode in which the rich combustion cycle and the lean combustion cycle are alternately switched in the same cylinder, that is, patterns 3 and 4 in FIG. Such a mode is selected (step 114).

  Subsequent to step 112 or 114, the magnitude of the negative valve overlap, the fuel injection amount, and the target value of the fuel injection timing in the rich combustion cycle are acquired in accordance with the rich combustion map described above (step 116). Further, the target value of the negative valve overlap, the fuel injection amount, and the fuel injection timing in the lean combustion cycle is acquired according to the above-described lean combustion map (step 118).

  When the processing of steps 116 and 118 described above has obtained the target values of the negative valve overlap magnitude, fuel injection amount, and fuel injection timing for each of the rich combustion cycle and the lean combustion cycle, The rich lean control is executed in the operation mode selected in step 112 or 114 using the target value (step 120).

  According to the processing of the routine shown in FIG. 8 described above, the same effect as in the first embodiment can be obtained. Furthermore, according to the routine processing shown in FIG. 8, it is possible to reliably avoid the in-cylinder temperature of the rich combustion cycle from becoming too high or too low when the rich lean control is executed. Therefore, it is possible to reliably prevent the smoke from being discharged and to suppress the occurrence of misfire and stabilize the combustion.

  In step 110, whether to select the rich combustion continuation mode or the rich / lean switching mode is determined according to the load, but this determination is made based on the exhaust temperature. Also good. That is, a determination value of the exhaust temperature is set, and when the exhaust temperature detected by the exhaust temperature sensor 47 exceeds the determination value, the rich combustion continuation mode is selected, and the detected exhaust temperature is equal to or lower than the determination value. In this case, the rich / lean switching mode may be selected. In this case, in a system in which the exhaust temperature sensor 47 is not provided, the exhaust temperature may be estimated from the operating state.

  In the second embodiment described above, the “mode switching means” according to the seventh aspect of the present invention is implemented when the ECU 50 executes the processes of steps 110 to 114.

Embodiment 3 FIG.
Next, the third embodiment of the present invention will be described with reference to FIG. 9. The description will focus on the differences from the first and second embodiments described above, and the description of the same matters will be omitted. Or simplify.

[Features of Embodiment 3]
FIG. 9 is a diagram illustrating an example of valve timing at the time of rich lean control in the third embodiment. As described in the second embodiment, in the rich combustion, it is important to prevent the in-cylinder temperature from becoming too high in order to prevent smoke emission. In this embodiment, when the rich lean control is executed, the intake valve closing timing (IVC) in the rich combustion cycle is set as shown in FIG. 9 in order to surely avoid that the in-cylinder temperature of the rich combustion cycle becomes too high. Therefore, it was decided to delay the intake valve closing timing in the lean combustion cycle.

  The later the intake valve closing timing is, the lower dead center is passed and the intake valve 52 is closed at the position where the piston 64 is raised. For this reason, as the intake valve closing timing is delayed, the gas once sucked into the cylinder is returned to the intake port 35, so that the amount of gas finally sucked into the cylinder can be reduced. . For this reason, even if the EGR amount is the same, the later the intake valve closing timing, the smaller the in-cylinder air amount (fresh air amount). Therefore, the later the intake valve closing timing is, the smaller the EGR amount (EGR rate) necessary for setting the air-fuel ratio A / F in the cylinder to the target air-fuel ratio.

  As described above, the in-cylinder temperature of the rich combustion cycle tends to increase as the EGR amount (EGR rate) increases. According to this embodiment, since the required EGR amount (EGR rate) can be reduced by delaying the intake valve closing timing of the rich combustion cycle, the in-cylinder temperature in the rich combustion cycle becomes too high. It can be effectively suppressed.

  Further, as the intake valve closing timing is delayed, the amount of gas finally sucked into the cylinder is reduced, so that the substantial compression ratio (hereinafter referred to as “actual compression ratio”) becomes smaller. When the actual compression ratio decreases, the compression end temperature, that is, the in-cylinder temperature at the start of combustion decreases. As a result, the combustion temperature also decreases.

  Thus, if the intake valve closing timing of the rich combustion cycle is delayed, the effect of reducing the required EGR amount (EGR rate) and the effect of lowering the combustion temperature combine to increase the in-cylinder temperature. Can be suppressed. For this reason, according to the present embodiment, it is possible to reliably avoid an excessive increase in the in-cylinder temperature of the rich combustion cycle, and to reliably prevent smoke discharge.

  Since Embodiment 3 is the same as Embodiment 1 or Embodiment 2 except for the points described above, further description thereof is omitted here.

  In the third embodiment described above, the intake variable valve mechanism 54 corresponds to the “variable valve mechanism” in the eighth invention. The ECU 50 controls the intake variable valve mechanism 54 so that the intake valve closing timing of the rich combustion cycle is later than the intake valve closing timing of the lean combustion cycle when the rich lean control is executed. "Combustion cycle intake valve slow closing means" is realized.

It is a figure for demonstrating the system configuration | structure of Embodiment 1 of this invention. It is a figure which shows the cross section of one cylinder of the diesel engine in the system shown in FIG. It is a figure for demonstrating a negative valve overlap. It is a figure which shows the relationship between the air fuel ratio A / F and the density | concentration of HC contained in the exhaust gas discharged | emitted by the combustion with the air fuel ratio A / F. It is a figure for demonstrating the example of the combustion mode of each cylinder in the case of performing rich lean control. It is a figure showing the change of generated torque at the time of changing the HC concentration in intake air, making the fuel injection quantity from an injector constant. It is a flowchart of the routine performed in Embodiment 1 of the present invention. It is a flowchart of the routine performed in Embodiment 2 of this invention. It is a figure which shows an example of the valve timing at the time of the rich lean control in Embodiment 3 of this invention.

Explanation of symbols

DESCRIPTION OF SYMBOLS 10 Diesel engine 12 Injector 14 Common rail 18 Exhaust passage 20 Exhaust manifold 22 Exhaust port 24 Turbo supercharger 26 Exhaust purification device 28 Intake passage 34 Intake manifold 36 Intake throttle valve 38 Air flow meter 40 External EGR passage 44 EGR valve 46 Intake pressure sensor 47 Exhaust temperature sensor 48 Accelerator opening sensor 50 ECU
52 Intake valve 54 Intake variable valve mechanism 56 Exhaust valve 58 Exhaust variable valve mechanism 62 Crank angle sensor 64 Piston

Claims (8)

An external EGR passage connecting an exhaust passage and an intake passage of a compression ignition type internal combustion engine;
An external EGR amount varying means for varying the amount of external EGR flowing back through the external EGR passage;
Internal EGR amount varying means for varying the amount of internal EGR generated inside the internal combustion engine without passing through the external EGR passage;
A lean combustion cycle that makes the air-fuel ratio leaner than the stoichiometric air-fuel ratio, and rich combustion that makes the air-fuel ratio relatively rich compared to the lean combustion cycle by making the EGR rate and fuel injection amount different from the lean combustion cycle Rich lean control means for performing rich lean control to switch cycles for each cycle in the same cylinder or to continue in separate cylinders,
With
The rich lean control means limits the external EGR rate during execution of the rich lean control to a predetermined upper limit value or less, and covers the remaining EGR amount necessary for each of the lean combustion cycle and the rich combustion cycle with the internal EGR. look including an EGR control means for controlling the state of the external EGR amount varying means and the internal EGR amount changing means as,
The rich combustion cycle is a cycle in which combustion is performed by reducing the in-cylinder temperature to a temperature at which smoke is not generated by increasing the EGR rate .   The predetermined upper limit value is the torque between the rich combustion cycle and the lean combustion cycle when the exhaust gas in the rich combustion cycle is recirculated by external EGR and mixed with the intake air so that the HC concentration in the intake air increases. 2. The control apparatus for an internal combustion engine according to claim 1, wherein the difference is a value determined so as to be within an allowable range.   2. The control apparatus for an internal combustion engine according to claim 1, wherein the EGR control means sets an external EGR rate during execution of the rich lean control to be substantially zero.   The internal EGR amount varying means includes a variable valve mechanism that varies the magnitude of positive or negative valve overlap between the intake valve and the exhaust valve for each cylinder, each cylinder group, or each cycle. The control device for an internal combustion engine according to any one of claims 1 to 3, wherein the control device is an internal combustion engine. An exhaust purification device disposed in the exhaust passage of the internal combustion engine;
The control apparatus for an internal combustion engine according to any one of claims 1 to 4, wherein the rich lean control means executes the rich lean control when regeneration of the exhaust purification device is required. .   In accordance with the load of the internal combustion engine, the rich lean control means is configured so that each of the EGR rate and the fuel injection amount is set so that the rich combustion cycle torque and the lean combustion cycle torque become equal in the low load region. And a torque control means for controlling each EGR rate and fuel injection amount so that the torque of the lean combustion cycle is larger than the torque of the rich combustion cycle in the high load side region. The control device for an internal combustion engine according to any one of claims 1 to 5.   The rich lean control means performs the lean combustion cycle and the rich combustion cycle in separate cylinders in a region of a high load side or a high exhaust gas temperature side according to a load or exhaust gas temperature of the internal combustion engine. And a mode switching means for switching the lean combustion cycle and the rich combustion cycle periodically in the same cylinder in the low load side or the low exhaust gas temperature side region. The control device for an internal combustion engine according to any one of claims 1 to 6. A variable valve mechanism that makes the intake valve closing timing variable;
When the rich lean control is executed, the rich valve that reduces the in-cylinder air amount in the rich combustion cycle by delaying the intake valve closing timing in the rich combustion cycle as compared with the intake valve closing timing in the lean combustion cycle. Combustion cycle intake valve slow closing means;
The control apparatus for an internal combustion engine according to any one of claims 1 to 7, further comprising:

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