JP2011089445A - Combustion control device of internal combustion engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a combustion control device of an internal combustion engine capable of inhibiting the generation of both NOx and smoke. <P>SOLUTION: At a low load operation of an engine, the injection of fuel for early slow combustion and premixed combustion is performed as first fuel injection at excessively advanced timing, and then second fuel injection for diffusion combustion is performed at a compression top dead point. Also, the injection amount of the fuel in the first fuel injection is increased as compared to that in the second fuel injection. An interference cooling action between injected each fuel is made to occur to contribute to a delay in ignition to increase the proportion of premixed combustion, thus resulting in the inhibition of the generation of smoke. By this, a large amount of EGR can be returned, whereby the generation of NOx can be reduced. <P>COPYRIGHT: (C)2011,JPO&INPIT

Description

  The present invention relates to a combustion control device for a compression ignition type internal combustion engine represented by a diesel engine. In particular, the present invention relates to measures for improving exhaust emission.

  In an engine that performs lean combustion, such as a diesel engine, the operating region that burns a mixture with a high air-fuel ratio (lean atmosphere) occupies most of the entire operating region, so nitrogen oxide (hereinafter referred to as NOx) There is a concern that a relatively large amount will be discharged. As a countermeasure, it is known to provide an exhaust gas recirculation (EGR) device that recirculates part of the exhaust gas to the intake passage (see, for example, Patent Document 1 and Patent Document 2 below). In other words, the exhaust gas is recirculated toward the cylinder to reduce the oxygen concentration and oxygen density in the cylinder. As a result, the combustion temperature (flame temperature) at the time of combustion in the combustion chamber is lowered to suppress the generation of NOx, thereby improving the exhaust emission.

  On the other hand, if incomplete combustion of the air-fuel mixture occurs during combustion in the combustion chamber, smoke is generated in the exhaust gas, leading to deterioration of exhaust emission. As a measure to reduce the amount of smoke generated, the main injection is divided into multiple stages, and the amount of fuel injected per fuel injection is reduced to eliminate oxygen shortages in the combustion field and reduce the amount of smoke generated. Suppression is done.

Japanese Patent Laid-Open No. 2004-3415 JP 2002-188487A JP 2004-3439 A

  By the way, when the diameter of the injection hole is reduced in order to atomize the spray injected from the injector for the purpose of improving combustibility, the fuel injection amount per unit time is reduced. For this reason, when the main injection is divided into multiple stages as described above during the low load operation of the engine in which the total fuel injection amount is set to be small, the injection amount in each divided injection also becomes small. The spraying starts in a state where the sprays are independent from each other in the combustion chamber.

  In such a situation, the temperature in the combustion chamber rapidly increases, and as a result, a diffusion combustion state is reached in the middle of multistage split injection, and a large amount of smoke is generated due to lack of oxygen in the combustion field. There is a possibility that.

  FIG. 10 shows the inside of the cylinder during the main injection execution period when three split injections (multi-stage split injection of the main injection) are performed after compression top dead center (ATDC) of the piston during low load operation of the engine. 2 shows an example of a change in heat generation rate and a fuel injection pattern.

  In the fuel injection pattern as shown in FIG. 10, the divided injections (particularly the second divided injection and the third divided injection) become independent combustions, respectively, and reach the diffusion combustion state at an early stage (ignition). It is in a state of insufficient delay), and a relatively large amount of smoke is generated due to insufficient oxygen in the combustion field. Even when the main injection is executed by two divided injections, if the first divided injection is executed at the timing after the compression top dead center of the piston, as in the case described above, a relatively large amount Smoke may occur.

  On the other hand, in order to suppress the generation amount of NOx as described above, it is effective to recirculate a part of the exhaust gas to the intake passage, but when the exhaust gas recirculation amount is increased to such a level that almost no NOx is generated. (For example, when the EGR rate is set to 20% or more), the in-cylinder temperature (the temperature of the combustion field before ignition) rises due to the heat of the exhaust gas that is recirculated. In other words, the temperature of the exhaust gas recirculated to the intake passage (hereinafter sometimes referred to as EGR gas) has a limit on the temperature that can be lowered even when cooling by the EGR cooler. In a situation where the intake system increases, the in-cylinder temperature significantly increases.

  In the situation where the amount of EGR gas recirculated in this way is increased, the diffusion combustion state resulting from the above-described split injection is started much earlier, so that a larger amount of smoke may be generated. There is sex.

  In consideration of the above points, the inventor of the present invention increases the premixed combustion amount in the combustion chamber in a state where the generation amount of NOx is suppressed by increasing the exhaust gas recirculation amount (the amount of fuel for performing premixed combustion). In this case, the inventors have focused on the fact that both the suppression of NOx generation and the suppression of smoke generation can be achieved.

  In Patent Document 3, the first set EGR value after the main injection is used in the main injection in the intake stroke or the compression stroke so that the premixed combustion ratio is higher than the diffusion combustion ratio. It is disclosed that the above EGR gas is controlled to recirculate. However, in this case, the EGR gas recirculation amount causes a transition from a cold flame reaction to a hot flame reaction at once, or a secondary flame that is executed at the end of the compression stroke causes the cold flame reaction to continue locally and cannot shift to a hot flame reaction. An area may occur. For this reason, premix combustion cannot be performed stably, and as a result, exhaust emission may be deteriorated.

  The present invention has been made in view of this point, and an object of the present invention is to provide a combustion control device for an internal combustion engine that can achieve both suppression of NOx generation amount and suppression of smoke generation amount.

-Principle of solving the problem-
The solution principle of the present invention devised to achieve the above object is that the ignition delay of the previously injected fuel among the divided main injections injected at a predetermined interval is determined as the final stage fuel. Each fuel injection timing is defined so as to promote the vicinity of the injection timing. As a result, a mutual interference cooling action between the fuel sprays is produced, and the proportion of premixed combustion during the combustion period is increased.

-Solution-
Specifically, the present invention includes an exhaust gas recirculation device that recirculates part of the exhaust gas discharged to the exhaust system to the intake system, and at least initially slows down as main injection that is fuel injection for generating torque. It is based on a combustion control device for a compression ignition type internal combustion engine equipped with a fuel injection valve configured to be capable of performing multiple stages of divided main injection including fuel injection for sequentially performing combustion and premixed combustion. . With respect to the combustion control device of the internal combustion engine, the divided main injection of the plurality of stages is performed in a state where a part of the exhaust gas is recirculated to the intake system by the exhaust gas recirculation device and the preheating operation before the start of the initial slow combustion is not performed. The final split main injection is executed in the vicinity of the compression top dead center of the piston, and the fuel injection that is performed before the final split main injection is executed on the advance side from the compression top dead center of the piston. Therefore, the timing of the heat generation rate from rising to lowering is delayed from the compression top dead center of the piston by combining the gas mass in the fuel injection executed on the advance side and the gas mass in the final split main injection. Fuel injection control means for performing fuel injection control is provided so that it exists at only one location on the corner side.

  As another solution, an exhaust gas recirculation device that recirculates a part of the exhaust gas discharged to the exhaust system to the intake system and at least an initial stage as main injection that is fuel injection for generating torque is provided. On the premise of a combustion control device for a compression self-ignition internal combustion engine having a fuel injection valve configured to be capable of performing multiple stages of divided main injection including fuel injection for performing slow combustion and premixed combustion in order To do. With respect to the combustion control device of the internal combustion engine, the divided main injection of the plurality of stages is performed in a state where a part of the exhaust gas is recirculated to the intake system by the exhaust gas recirculation device and the preheating operation before the start of the initial slow combustion is not performed The final split main injection is executed in the vicinity of the compression top dead center of the piston, and the fuel injection that is performed before the final split main injection is executed on the advance side from the compression top dead center of the piston. Thus, there is provided fuel injection control means for executing fuel injection control for causing an interference cooling action by the gas mass in the final divided main injection with respect to the gas mass in the fuel injection executed on the advance side.

  Due to these specific matters, the fuel injected into the cylinder by the fuel injection (hereinafter referred to as the first fuel injection) performed before the final divided main injection is advanced from the compression top dead center of the piston. As a result of being supplied into the combustion chamber, an overdiffusion state occurs in the cylinder, resulting in an ignition delay. On the other hand, the fuel injected into the cylinder by the final split main injection (hereinafter referred to as second fuel injection) is supplied into the combustion chamber near the compression top dead center of the piston, so that the ignition delay in the cylinder is reduced. It will hardly occur. Due to the difference in the ignition delay amount, the fuel spray (gas lump) injected by the first fuel injection and the fuel spray (gas lump) injected by the second fuel injection are combined. At this time, the spray injected by the second fuel injection undergoes an endothermic reaction, thereby cooling the spray injected by the first fuel injection. In this way, after fuel injection in the second fuel injection, the temperature in the combustion chamber is kept low by the mutual interference cooling action in the combustion chamber, and each spray (spray in the first fuel injection and the second fuel). The spraying by injection) is combined, and the initial slow combustion and the premixed combustion are sequentially performed. Therefore, the timing at which the heat generation rate changes from rising to falling (peak timing of the heat generation rate) is present at only one location on the retard side from the compression top dead center of the piston. As a result, most of the combustion modes of the fuel injected in the first fuel injection and the second fuel injection are premixed combustion, and oxygen shortage does not occur in the combustion field, so that the generation of smoke is suppressed. Can do. In addition, since the generation of smoke can be suppressed in this manner, the EGR rate can be set high, and the amount of NOx generated can be greatly reduced. In the present solution, the first fuel injection and the second fuel injection are not limited to one fuel injection each, and one or both of the fuel injections may be further divided.

  The period of the final divided main injection is preferably set as a period from before the compression top dead center to after the compression top dead center.

  According to this, during the fuel injection period of the second fuel injection (final divided main injection), the volume of the combustion chamber is minimized, and the second fuel injection is performed with the combustion chamber temperature being relatively high. Therefore, the fuel ignition delay in the second fuel injection is substantially minimized. As a result, the difference in ignition delay between the fuel injected by the first fuel injection and the fuel injected by the second fuel injection can be set to the maximum, and the above-described mutual interference cooling action can be effectively performed. It can be demonstrated.

  In order to switch the mode of the cooling action in the combustion chamber according to the load state of the internal combustion engine, the following configuration is preferable. That is, during low-load operation of the internal combustion engine, fuel injection that is performed at a stage prior to the final split main injection is executed on the advance side of the compression top dead center of the piston, and the final split main injection is On the other hand, during high-load operation of the internal combustion engine, fuel injection performed before the final divided main injection is performed near the compression top dead center of the piston, and the final divided main injection is performed on the piston. The configuration is such that it is executed on the retard side from the compression top dead center.

  In this case, at the time of low load operation of the internal combustion engine, combustion is performed while the temperature in the combustion chamber is kept low by the mutual interference cooling action between the first fuel injection and the second fuel injection, and the amount of smoke generated and NOx is reduced. Can be suppressed. On the other hand, during the high load operation of the internal combustion engine, most of the injection period of the fuel injection (first fuel injection) performed before the final divided main injection is preceded and injected into the combustion chamber. The fuel is then cooled by the endothermic reaction of the fuel subsequently injected into the combustion chamber. By continuing such a cooling action (self-interference cooling action) by the subsequent fuel, an increase in the combustion field temperature (temperature in the region where the spray exists) in the combustion chamber is suppressed. For this reason, most of the combustion mode of the fuel injected during this injection period is premixed combustion, and oxygen shortage does not occur in the combustion field, so that the generation of smoke can be suppressed. Further, since the generation of smoke can be suppressed by such a self-interference cooling action, the EGR rate by the exhaust gas recirculation device can be set high, and the generation amount of NOx can be greatly reduced. Thus, according to this solution, it is possible to achieve both the suppression of the NOx generation amount and the suppression of the smoke generation amount by generating a cooling action corresponding to any load state.

  As a configuration for setting the fuel injection amount in each fuel injection period, the following is preferable. In other words, the fuel injection control means sets the penetration force of the fuel injected in the final divided main injection lower than the penetration force of the fuel injected in the fuel injection performed in the stage before the final divided main injection. Is configured. More specifically, the total fuel injection amount injected in the final divided main injection is set to be smaller than the total fuel injection amount injected in the fuel injection performed in a stage before the final divided main injection. The fuel injection control means is configured.

  Thus, by setting the fuel injection amount in each fuel injection period and making each penetration force different, combustion in a combustion field over a wide range in the combustion chamber becomes possible, and oxygen shortage in each combustion field is reduced. This can be avoided, and the generation of smoke can be more reliably suppressed.

  More preferably, the injection timing and the injection amount of the fuel injection performed before the final divided main injection so that substantially the entire fuel injected by the final divided main injection becomes initial slow combustion and premixed combustion. Set. For that purpose, for example, the injection timing of the first fuel injection is largely shifted to the advance side, and the fuel injected in the first fuel injection is overdiffusioned so that no preheating action occurs. It is done.

  According to this, the combustion mode in the combustion period can be achieved only by the initial slow combustion and the premixed combustion. That is, since there is no diffusion combustion, it is possible to reliably achieve both suppression of the NOx generation amount and suppression of the smoke generation amount.

  The following operation is preferably performed as an auxiliary operation for suppressing the generation of smoke and NOx. That is, fuel injection performed before the final divided main injection is performed while performing a drought reduction operation for reducing the droop rate between oxygen in the cylinder and fuel spray.

  Here, the “rate of oxygen and fuel spray in the cylinder” is a so-called “meeting (performing chemical reaction)” probability (frequency) of oxygen molecules and fuel particles in the cylinder. The higher the rate, the more the chemical reaction in the cylinder progresses, and the in-cylinder temperature rises as the heat generation rate increases. In other words, if the operation for reducing the porosity is performed as described above, even if the amount of oxygen molecules is small or the amount of fuel particles is large in a partial region in the cylinder (for example, a narrow region in the center of the combustion chamber). Even so, the degree of progress of the chemical reaction is low, and the temperature inside the cylinder can be lowered (for example, about 800 K). For example, the initial slow combustion can be realized by suppressing the number of collisions between oxygen molecules and fuel particles per unit volume in the cylinder or suppressing the momentum of oxygen molecules.

  Thus, since fuel injection is performed in a state where the ratio of oxygen and fuel spray in the cylinder is reduced, the heat generation rate, which is the amount of heat generated per unit time, is relatively small. Occurrence can be suppressed.

  Preferably, at least one of the exhaust gas recirculation operation by the exhaust gas recirculation device, the intake air throttle operation in the intake system, and the operation of lowering the in-cylinder temperature is executed as the ratio reduction operation.

  By these operations, the above-described ratio is effectively reduced by lowering the oxygen concentration in the intake air, reducing the intake air amount, or reducing the kinetic energy of oxygen molecules and fuel particles in the cylinder. be able to.

  The exhaust gas recirculation rate in a state where a part of the exhaust gas is recirculated to the intake system by the exhaust gas recirculation device is preferably set to 20% or more. By setting the exhaust gas recirculation rate to a relatively high value in this way, it is possible to greatly reduce the amount of NOx generated.

  Furthermore, it is preferable that the fuel injection valve has a nozzle hole diameter reduced so that the fuel injection rate when the valve is opened is equal to or less than a predetermined injection rate.

  When the nozzle hole diameter is reduced in this way, the combustibility is improved by atomizing the spray, and for example, the amount of HC generated can be suppressed. However, there is a possibility of prolonging the fuel injection period. In this case, by performing the fuel injection control so as to obtain the mutual interference cooling action as described above, it is possible to increase the proportion of the premixed combustion by promoting the ignition delay, thereby suppressing the NOx generation amount and the smoke generation amount. It is possible to achieve both suppression of the above.

  In the present invention, each fuel injection timing is defined so as to accelerate the ignition delay of the previously injected fuel to the vicinity of the final stage fuel injection timing, thereby causing the mutual interference cooling action between the fuel sprays, and the combustion period. The proportion of premixed combustion in the engine is increased. As a result, it is possible to achieve both the suppression of the NOx generation amount and the suppression of the smoke generation amount, and the exhaust emission can be improved.

It is a schematic block diagram of the engine which concerns on embodiment, and its control system. It is sectional drawing which shows the combustion chamber of a diesel engine, and its peripheral part. It is a block diagram which shows the structure of control systems, such as ECU. It is a figure which shows the change of the heat release rate in a cylinder at the time of a low load driving | operation of an engine, and a fuel-injection pattern, respectively. It is a figure which shows the change of the heat release rate in a cylinder at the time of a high load driving | operation of an engine, and a fuel-injection pattern, respectively. FIG. 6 is a φT map showing changes in gas temperature and equivalence ratio of a combustion field when each divided main injection is performed, and shows changes during low-load operation of the engine. FIG. 5 is a φT map showing changes in the gas temperature of the combustion field and the equivalence ratio when each divided main injection is performed, and is a diagram showing changes during high-load operation of the engine. It is a flowchart figure which shows the procedure of fuel injection form change control. It is a modification, and is a diagram showing a change in a heat generation rate in a cylinder and a fuel injection pattern at the time of low load operation of the engine. In a prior art example, it is a figure which respectively shows the change of the heat release rate in a cylinder in the execution period of the main injection at the time of three times of divided injection, and a fuel injection pattern.

  Hereinafter, embodiments of the present invention will be described with reference to the drawings. In the present embodiment, a case where the present invention is applied to a common rail in-cylinder direct injection multi-cylinder (for example, in-line 4-cylinder) diesel engine (compression self-ignition internal combustion engine) mounted on an automobile will be described.

-Engine configuration-
First, a schematic configuration of a diesel engine (hereinafter simply referred to as an engine) according to the present embodiment will be described. FIG. 1 is a schematic configuration diagram of an engine 1 and its control system according to the present embodiment. FIG. 2 is a cross-sectional view showing the combustion chamber 3 of the diesel engine and its periphery.

  As shown in FIG. 1, the engine 1 according to the present embodiment is configured as a diesel engine system having a fuel supply system 2, a combustion chamber 3, an intake system 6, an exhaust system 7 and the like as main parts.

  The fuel supply system 2 includes a supply pump 21, a common rail 22, an injector (fuel injection valve) 23, a shutoff valve 24, a fuel addition valve 26, an engine fuel passage 27, an addition fuel passage 28, and the like.

  The supply pump 21 pumps fuel from the fuel tank, makes the pumped fuel high pressure, and supplies it to the common rail 22 via the engine fuel passage 27. The common rail 22 has a function as a pressure accumulation chamber that holds (accumulates) the high-pressure fuel supplied from the supply pump 21 at a predetermined pressure, and distributes the accumulated fuel to the injectors 23. The injector 23 includes a piezoelectric element (piezo element) therein, and is configured by a piezo injector that is appropriately opened to supply fuel into the combustion chamber 3.

  Moreover, the injector 23 in this embodiment is provided with ten nozzle holes (not shown) at equal intervals in the circumferential direction, and each nozzle hole diameter is made smaller than that of a general injector. . This is a configuration for improving combustibility by atomizing the spray and suppressing, for example, the amount of HC generated. In such a configuration, the fuel injection amount (fuel injection rate) per unit time when the injector 23 is opened is small. Specifically, the nozzle hole diameter of a general injector is 0.12 mm, whereas the nozzle hole diameter of the injector 23 in this embodiment is 0.09 mm. This value is not limited to this. Details of the fuel injection control from the injector 23 will be described later.

  The supply pump 21 supplies a part of the fuel pumped from the fuel tank to the fuel addition valve 26 via the addition fuel passage 28. The added fuel passage 28 is provided with the shutoff valve 24 for shutting off the added fuel passage 28 and stopping fuel addition in an emergency.

  The fuel addition valve 26 is configured so that the fuel addition amount to the exhaust system 7 becomes a target addition amount (addition amount that makes the exhaust A / F become the target A / F) by an addition control operation by the ECU 100 described later. In addition, it is constituted by an electronically controlled on-off valve whose valve opening timing is controlled so that the fuel addition timing becomes a predetermined timing. That is, a desired fuel is injected and supplied from the fuel addition valve 26 to the exhaust system 7 (from the exhaust port 71 to the exhaust manifold 72) at an appropriate timing.

  The intake system 6 includes an intake manifold 63 connected to an intake port 15a formed in the cylinder head 15 (see FIG. 2), and an intake pipe 64 that constitutes an intake passage is connected to the intake manifold 63. Further, an air cleaner 65, an air flow meter 43, and a throttle valve (intake throttle valve) 62 are arranged in this intake passage in order from the upstream side. The air flow meter 43 outputs an electrical signal corresponding to the amount of air flowing into the intake passage via the air cleaner 65.

  The exhaust system 7 includes an exhaust manifold 72 connected to an exhaust port 71 formed in the cylinder head 15, and exhaust pipes 73 and 74 constituting an exhaust passage are connected to the exhaust manifold 72. In addition, a maniverter (exhaust gas purification device) 77 including a NOx storage catalyst (NSR catalyst: NOx Storage Reduction catalyst) 75 and a DPNR catalyst (Diesel Particle-NOx Reduction catalyst) 76 is disposed in the exhaust passage. Hereinafter, the NSR catalyst 75 and the DPNR catalyst 76 will be described.

The NSR catalyst 75 is an NOx storage reduction catalyst. For example, alumina (Al ₂ O ₃ ) is used as a support, and potassium (K), sodium (Na), lithium (Li), cesium (Cs), for example, is supported on this support. Alkali metal such as barium (Ba), alkaline earth such as calcium (Ca), rare earth such as lanthanum (La) and yttrium (Y), and noble metal such as platinum (Pt) were supported. It has a configuration.

The NSR catalyst 75 stores NOx in a state where a large amount of oxygen is present in the exhaust gas, has a low oxygen concentration in the exhaust gas, and has a large amount of reducing component (for example, unburned component (HC) of fuel). In the existing state, NOx is reduced to NO ₂ or NO and released. NO NOx released as NO ₂ or NO, the N ₂ is further reduced due to quickly reacting with HC or CO in the exhaust. Further, HC and CO are oxidized to H ₂ O and CO ₂ by reducing NO ₂ and NO. That is, HC, CO, and NOx in the exhaust can be purified by appropriately adjusting the oxygen concentration and HC component in the exhaust introduced into the NSR catalyst 75. In the present embodiment, the oxygen concentration and HC component in the exhaust gas can be adjusted by the fuel addition operation from the fuel addition valve 26.

  On the other hand, the DPNR catalyst 76 is, for example, a porous ceramic structure carrying a NOx storage reduction catalyst, and PM in the exhaust gas is collected when passing through the porous wall. Further, when the air-fuel ratio of the exhaust gas is lean, NOx in the exhaust gas is stored in the NOx storage reduction catalyst, and when the air-fuel ratio becomes rich, the stored NOx is reduced and released. Further, the DPNR catalyst 76 carries a catalyst that oxidizes and burns the collected PM (for example, an oxidation catalyst mainly composed of a noble metal such as platinum).

  Here, the structure of the combustion chamber 3 of a diesel engine and its peripheral part is demonstrated using FIG. As shown in FIG. 2, a cylinder block 11 constituting a part of the engine body is formed with a cylindrical cylinder bore 12 for each cylinder (four cylinders), and a piston 13 is formed inside each cylinder bore 12. Is accommodated so as to be slidable in the vertical direction.

  The combustion chamber 3 is formed above the top surface 13 a of the piston 13. That is, the combustion chamber 3 is defined by the lower surface of the cylinder head 15 attached to the upper part of the cylinder block 11 via the gasket 14, the inner wall surface of the cylinder bore 12, and the top surface 13 a of the piston 13. A cavity (concave portion) 13 b is formed in a substantially central portion of the top surface 13 a of the piston 13, and this cavity 13 b also constitutes a part of the combustion chamber 3.

  As for the shape of the cavity 13b, the concave dimension is small in the central portion (on the cylinder center line P), and the concave dimension is increased toward the outer peripheral side. That is, as shown in FIG. 2, when the piston 13 is in the vicinity of the compression top dead center, the combustion chamber 3 formed by the cavity 13b is a narrow space having a relatively small volume at the center portion, and is directed toward the outer peripheral side. Thus, the space is gradually enlarged (expanded space).

  The piston 13 has a small end portion 18a of a connecting rod 18 connected by a piston pin 13c, and a large end portion of the connecting rod 18 is connected to a crankshaft which is an engine output shaft. As a result, the reciprocating movement of the piston 13 in the cylinder bore 12 is transmitted to the crankshaft via the connecting rod 18, and the engine output is obtained by rotating the crankshaft. Further, a glow plug 19 is disposed toward the combustion chamber 3. The glow plug 19 functions as a start-up assisting device that is heated red when an electric current is applied immediately before the engine 1 is started and a part of the fuel spray is blown onto the glow plug 19 to promote ignition and combustion.

  The cylinder head 15 is formed with an intake port 15a for introducing air into the combustion chamber 3 and an exhaust port 71 for discharging exhaust gas from the combustion chamber 3, and an intake valve for opening and closing the intake port 15a. 16 and an exhaust valve 17 for opening and closing the exhaust port 71 are provided. The intake valve 16 and the exhaust valve 17 are disposed to face each other with the cylinder center line P interposed therebetween. That is, the engine 1 is configured as a cross flow type. The cylinder head 15 is provided with the injector 23 that directly injects fuel into the combustion chamber 3. The injector 23 is disposed at a substantially upper center of the combustion chamber 3 in a standing posture along the cylinder center line P, and injects fuel introduced from the common rail 22 toward the combustion chamber 3 at a predetermined timing. It has become.

  Furthermore, as shown in FIG. 1, the engine 1 is provided with a supercharger (turbocharger) 5. The turbocharger 5 includes a turbine wheel 52 and a compressor wheel 53 that are connected via a turbine shaft 51. The compressor wheel 53 is disposed facing the inside of the intake pipe 64, and the turbine wheel 52 is disposed facing the inside of the exhaust pipe 73. For this reason, the turbocharger 5 performs a so-called supercharging operation in which the compressor wheel 53 is rotated using the exhaust flow (exhaust pressure) received by the turbine wheel 52 to increase the intake pressure. The turbocharger 5 in the present embodiment is a variable nozzle type turbocharger, and a variable nozzle vane mechanism (not shown) is provided on the turbine wheel 52 side. By adjusting the opening of the variable nozzle vane mechanism, the engine 1 supercharging pressure can be adjusted.

  An intake pipe 64 of the intake system 6 is provided with an intercooler 61 for forcibly cooling the intake air whose temperature has been raised by supercharging in the turbocharger 5. The throttle valve 62 provided further downstream than the intercooler 61 is an electronically controlled on-off valve whose opening degree can be adjusted steplessly. It has a function of narrowing down the area and adjusting (reducing) the supply amount of the intake air.

  Further, the engine 1 is provided with an exhaust gas recirculation passage (EGR passage) 8 that connects the intake system 6 and the exhaust system 7. The EGR passage 8 is configured to reduce the combustion temperature by recirculating a part of the exhaust gas to the intake system 6 and supplying it again to the combustion chamber 3, thereby reducing the amount of NOx generated. In addition, the EGR passage 8 is opened and closed steplessly by electronic control, and the exhaust gas passing through the EGR passage 8 (recirculating) is cooled by an EGR valve 81 that can freely adjust the exhaust flow rate flowing through the passage. An EGR cooler 82 is provided. The EGR passage 8, the EGR valve 81, the EGR cooler 82, and the like constitute an EGR device (exhaust gas recirculation device).

-Sensors-
Various sensors are attached to each part of the engine 1, and signals related to the environmental conditions of each part and the operating state of the engine 1 are output.

  For example, the air flow meter 43 outputs a detection signal corresponding to the flow rate of intake air (intake air amount) upstream of the throttle valve 62 in the intake system 6. The intake air temperature sensor 49 is disposed in the intake manifold 63 and outputs a detection signal corresponding to the temperature of the intake air. The intake pressure sensor 48 is disposed in the intake manifold 63 and outputs a detection signal corresponding to the intake air pressure. The A / F (air-fuel ratio) sensor 44 outputs a detection signal that continuously changes in accordance with the oxygen concentration in the exhaust gas downstream of the manipulator 77 of the exhaust system 7. Similarly, the exhaust temperature sensor 45 outputs a detection signal corresponding to the temperature of the exhaust gas (exhaust temperature) downstream of the manipulator 77 of the exhaust system 7. The rail pressure sensor 41 outputs a detection signal corresponding to the fuel pressure stored in the common rail 22. The throttle opening sensor 42 detects the opening of the throttle valve 62.

-ECU-
As shown in FIG. 3, the ECU 100 includes a CPU 101, a ROM 102, a RAM 103, a backup RAM 104, and the like. The ROM 102 stores various control programs, maps that are referred to when the various control programs are executed, and the like. The CPU 101 executes various arithmetic processes based on various control programs and maps stored in the ROM 102. The RAM 103 is a memory that temporarily stores calculation results in the CPU 101, data input from each sensor, and the like. The backup RAM 104 is a non-volatile memory that stores data to be saved when the engine 1 is stopped, for example.

  The CPU 101, the ROM 102, the RAM 103, and the backup RAM 104 are connected to each other via the bus 107, and are connected to the input interface 105 and the output interface 106.

  The input interface 105 is connected with the rail pressure sensor 41, the throttle opening sensor 42, the air flow meter 43, the A / F sensor 44, the exhaust temperature sensor 45, the intake pressure sensor 48, and the intake temperature sensor 49. Further, the input interface 105 includes a water temperature sensor 46 that outputs a detection signal corresponding to the cooling water temperature of the engine 1, an accelerator opening sensor 47 that outputs a detection signal corresponding to the depression amount of the accelerator pedal, and the engine 1. A crank position sensor 40 that outputs a detection signal (pulse) each time the output shaft (crankshaft) rotates by a certain angle is connected. On the other hand, the injector 23, the fuel addition valve 26, the throttle valve 62, the EGR valve 81, and the like are connected to the output interface 106.

  The ECU 100 executes various controls of the engine 1 based on the outputs of the various sensors described above. For example, the ECU 100 controls the opening degree of the EGR valve 81 according to the operating state of the engine 1 and adjusts the exhaust gas recirculation amount (EGR amount) toward the intake manifold 63. The EGR amount is set according to an EGR map stored in advance in the ROM 102. Specifically, this EGR map is a map for determining the EGR amount (EGR rate) using the engine speed and the engine load as parameters, and is for setting an EGR amount that can reduce the NOx emission amount to the exhaust system. belongs to. This EGR map is created in advance by experiments, simulations, or the like. That is, by applying the engine speed calculated based on the detection value of the crank position sensor 40 and the opening of the throttle valve 62 (corresponding to the engine load) detected by the throttle opening sensor 42 to the EGR map. An EGR amount (opening degree of the EGR valve 81) is obtained. In particular, the engine 1 according to this embodiment has a relatively large EGR amount. Specifically, the EGR map is created and stored in the ROM 102 so that the opening degree of the EGR valve 81 that always has an EGR rate of 20% or more is obtained regardless of the load state of the engine 1.

  Further, the ECU 100 executes fuel injection control for the injector 23. As fuel injection control of the injector 23, in this embodiment, sub-injection such as pilot injection, pre-injection, after-injection, and post-injection that is executed in a conventional general diesel engine is not executed, and engine torque is obtained. Therefore, only the main injection is performed. That is, by not performing the pilot injection or the pre-injection, the preheating operation until the main injection is started is not performed.

  The total fuel injection amount in the main injection is a fuel injection amount necessary for obtaining a required torque that is determined according to an operating state such as engine speed, accelerator operation amount, cooling water temperature, intake air temperature, and environmental conditions. Is set. For example, the higher the engine speed (the engine speed calculated based on the detection value of the crank position sensor 40), the larger the accelerator operation amount (the accelerator pedal depression amount detected by the accelerator opening sensor 47). The higher the required accelerator torque of the engine 1, the higher the accelerator opening.

-Fuel injection pressure-
The fuel injection pressure for executing the main fuel injection is determined by the internal pressure of the common rail 22. As the common rail internal pressure, generally, the target value of the fuel pressure supplied from the common rail 22 to the injector 23, that is, the target rail pressure, increases as the engine load (engine load) increases and the engine speed (engine speed) increases. It will be expensive. That is, when the engine load is high, the amount of air sucked into the combustion chamber 3 is large. Therefore, a large amount of fuel must be injected from the injector 23 into the combustion chamber 3, and therefore the injection from the injector 23 is performed. The pressure needs to be high. Further, when the engine speed is high, the injection period is short, so the amount of fuel injected per unit time must be increased, and therefore the injection pressure from the injector 23 needs to be increased. . Thus, the target rail pressure is generally set based on the engine load and the engine speed. The target rail pressure is set according to a fuel pressure setting map stored in the ROM 102, for example. That is, by determining the fuel pressure according to this fuel pressure setting map, the valve opening period (injection rate waveform) of the injector 23 is controlled, and the fuel injection amount during the valve opening period can be defined. In the present embodiment, the fuel pressure is adjusted between 30 MPa and 200 MPa according to the engine load and the like.

  The optimum value of the fuel injection parameter in the main injection varies depending on the temperature conditions of the engine 1 and the intake air.

  For example, the ECU 100 adjusts the fuel discharge amount of the supply pump 21 so that the common rail pressure becomes equal to the target rail pressure set based on the engine operating state, that is, the fuel injection pressure matches the target injection pressure. To measure. Further, the ECU 100 determines the fuel injection amount and the fuel injection form based on the engine operating state. Specifically, the ECU 100 calculates the engine rotation speed based on the detection value of the crank position sensor 40, obtains the amount of depression of the accelerator pedal (accelerator opening) based on the detection value of the accelerator opening sensor 47, A total main injection amount (injection amount in main injection) is determined based on the engine speed and the accelerator opening.

-Split main injection-
In the diesel engine 1, it is important to simultaneously satisfy various requirements such as improvement of exhaust emission by reducing the NOx generation amount and smoke generation amount, and reduction of combustion noise during the combustion stroke. The inventor of the present invention pays attention to the fact that it is effective to appropriately control the combustion mode in the combustion chamber 3 as a method for simultaneously satisfying these requirements, and as a method for controlling this combustion mode. The present inventors have found a fuel injection method using divided main injection (divided main injection) as described below. This will be specifically described below.

  In this embodiment, the injection mode of the main injection is changed according to the engine load. Specifically, in the fuel injection control in the present embodiment, regardless of the load of the engine 1 (during low load operation, medium load operation, and high load operation), the main injection mode is divided into two divided mains. By executing the injection, the total main injection amount (total fuel injection amount for obtaining the required torque) required for the main injection is secured.

  Note that the high load operation in the present embodiment means a region where the engine load is relatively high within the range of the operation region where exhaust emission is improved. That is, this is a region where the engine load is lower than the operation region (torque priority operation region) in which the torque requested by the driver is output without considering the exhaust emission.

  The divided main injection, which is a feature of the present embodiment, is at the injection timing of the divided main injection during the low load operation of the engine 1. Details of the injection mode at each load will be described later.

  The outline of the combustion mode during the combustion process in the combustion chamber 3 in the present embodiment is as follows. In this combustion mode, initial slow combustion (also referred to as initial low temperature combustion, which is a combustion mode having a relatively low combustion rate), premixed combustion, and diffusion combustion are sequentially performed during the same combustion process. In other words, during the same combustion process, the initial slow combustion and the diffusion combustion are performed, and furthermore, by connecting the initial slow combustion and the diffusion combustion by the premixed combustion, three different forms of combustion are continuously performed. It is supposed to be done. That is, by interposing the premixed combustion that is the second-stage combustion between the initial slow combustion that is the first-stage combustion and the diffusion combustion that is the third-stage combustion, a series of these different combustion forms Realize the sex.

  More specifically, the initial slow combustion, which is the first-stage combustion, is performed by injecting fuel while performing a drought reduction operation for reducing the droop rate between oxygen in the cylinder and fuel spray. The fuel burns sequentially in the cylinder as the fuel burns. Specifically, a predetermined period when the in-cylinder temperature is in the range of 750 K or more and less than 900 K is set as the initial combustion fuel injection period, and the slow combustion main injection is performed to perform the initial slow combustion. The upper limit value of the in-cylinder temperature during the initial combustion fuel injection period is a value that can be appropriately set according to the spray state in the cylinder, and may be increased by a large amount of EGR. This value is not limited to 900K, and may be 950K, 1000K, or the like (hereinafter, the case of 900K will be described as a representative).

  Premixed combustion, which is the second stage combustion, is performed when the temperature in the cylinder is lower than a predetermined diffusion combustion start temperature (for example, 900 K) leading to diffusion combustion after the start of the initial slow combustion. This is done by burning the existing fuel. Specifically, a predetermined period when the combustion field temperature of the initial slow combustion is in a range of 800 K or more and less than 900 K is set as a premixed combustion fuel injection period, the premixed combustion main injection is performed, and the premixed combustion is performed. Perform mixed combustion.

  In the present embodiment, as will be described later, in the middle load operation and the high load operation, the initial combustion fuel injection period for performing the initial slow combustion and the premixed combustion for performing the premixed combustion are performed. Each fuel injection period is set as a continuous fuel injection period. In other words, the fuel injection in the initial combustion fuel injection period for performing the initial slow combustion and the fuel injection in the premixed combustion fuel injection period for performing the premixed combustion do not stop the fuel injection. It is intended to be continued (continuous). Further, as will be described later, when the engine 1 is in a low load operation, the fuel injection timing is made different between the middle load operation and the high load operation, and the second divided main injection ( The initial slow combustion and the premixed combustion are sequentially performed in accordance with the execution of the divided main injection set near the compression top dead center (TDC) of the piston 13. Details will be described later.

  Further, diffusion combustion, which is the third stage combustion, is combustion by fuel injection when the temperature in the cylinder is equal to or higher than the diffusion combustion start temperature by the premix combustion after the start of the premix combustion. is there. Specifically, the diffusion combustion is performed by setting the predetermined period immediately after the combustion field temperature of the premixed combustion reaches 900K as the diffusion combustion fuel injection period and performing the diffusion combustion main injection. At the time of medium load operation and high load operation of the engine 1, the diffusion combustion fuel injection period is set on the retard side (ATDC) from the compression top dead center of the piston 13. On the other hand, during the low load operation of the engine 1, diffusion combustion is performed after the initial slow combustion and the premixed combustion are sequentially performed in accordance with the execution of the second divided main injection among the divided main injections. It will be.

  In the present embodiment, since the split main injection is executed twice as the main injection mode, the first split main injection is hereinafter referred to as “first fuel injection”, and the second split main injection is referred to as the main injection mode. This is referred to as “second fuel injection (final divided main injection)”.

  As the respective fuel injection amounts for realizing each combustion mode, when the engine 1 is in a low load operation, the fuel injection amount in the second fuel injection is less than the fuel injection amount in the first fuel injection. It is set. Thus, by setting the fuel injection amount at the time of low load operation, the penetration force of the fuel injected by the second fuel injection is set lower than the penetration force of the fuel injected by the first fuel injection. . The ratio of the fuel injection amount in the first fuel injection and the fuel injection amount in the second fuel injection is set to “2: 1”, for example. This value is not limited to this.

  Further, even during the medium load operation and the high load operation of the engine 1, the fuel injection amount in the second fuel injection is set smaller than the fuel injection amount in the first fuel injection. Thus, by setting the fuel injection amount at the time of high load operation, the penetration force of the fuel injected by the second fuel injection is set lower than the penetration force of the fuel injected by the first fuel injection. Yes. The ratio of the fuel injection amount in the first fuel injection and the fuel injection amount in the second fuel injection is set to “7: 1”, for example. This value is not limited to this.

  Here, the relationship between the fuel injection amount and the penetration force will be described. In the injector 23, when fuel injection is started in response to the injection command signal, the opening area of the injection hole is gradually increased as the needle closing the injection hole is retracted from the injection hole. When the needle moves to the last retracted position, the opening area of the injection hole becomes maximum. However, if the injection command signal is canceled before the needle reaches the last retracted position (when the valve closing command is received), the needle moves forward in the valve closing direction while moving backward. That is, in this case, the fuel injection is terminated without maximizing the opening area of the injection hole. For this reason, the longer the injection period is set, the larger the opening area of the injection hole is obtained.

  And the opening area of the said injection hole has a correlation with the flight distance of the fuel (spray) injected from the injection hole. In other words, when fuel is injected with a large opening area of the injection hole, the size of the droplet of fuel injected from the injection hole is also large, so that the kinetic energy is also large (the penetration force is large). Yes. For this reason, the flight distance of this fuel droplet becomes long. On the other hand, when fuel is injected in a state where the opening area of the injection hole is small, the size of the droplet of fuel injected from the injection hole is also small, so the kinetic energy is also small (penetration force is small). ing. For this reason, the flight distance of this fuel droplet is also short.

  As described above, when the valve opening period of the injector 23 is set to be relatively long (in other words, when the injection amount per main injection is set to be relatively large), the needle reaches the last retracted position. Since the opening area of the injection hole becomes maximum due to the movement, the flight distance of the fuel droplet in this case becomes long. That is, most of the fuel injected from the injector 23 can fly up to the vicinity of the outer peripheral end of the cavity 13b.

  On the other hand, if the valve opening period of the injector 23 is set to be relatively short (in other words, the injection amount per main injection is set to be relatively small), the needle may move to the last retracted position. Since the opening area of the injection hole is small, the flight distance of the fuel droplet in this case is shortened. That is, most of the fuel injected from the injector 23 can fly only to the vicinity of the center of the cavity 13b.

  Thus, there is a correlation between the opening area of the injection hole determined by the valve opening period of the injector 23 and the flight distance of the fuel (spray) injected from the injection hole. For this reason, it is possible to adjust the flight distance of the fuel by adjusting the valve opening period of the injector 23. In other words, there is a correlation between the opening area of the injection hole determined by the injection amount per main injection and the flight distance of fuel (spray) injected from the injection hole. For this reason, it is possible to define the flight distance of the fuel by defining the injection amount per main injection.

  Hereinafter, the fuel injection mode in each load state and the combustion mode in the combustion chamber 3 associated therewith will be individually described.

  FIG. 4 shows a change in the heat generation rate in the cylinder and the fuel injection pattern during the low-load operation of the engine 1 and during the main injection execution period. FIG. 5 shows the change in the heat generation rate in the cylinder and the fuel injection pattern during the high-load operation of the engine 1 and during the execution period of the main injection.

  4 and 5, the horizontal axis indicates the crank angle and the vertical axis indicates the heat generation rate. Further, in the waveform of the fuel injection pattern in each figure, the horizontal axis represents the crank angle, and the vertical axis represents the injection rate (corresponding to the backward movement amount of the needle provided in the injector 23). TDC in the figure indicates the crank angle position corresponding to the compression top dead center of the piston 13.

  Note that the change in the heat generation rate in the cylinder during the medium load operation and the fuel injection pattern are set to a small amount of fuel injection compared to that during the high load operation (the first fuel injection and the second fuel injection). And the injection timing thereof is the same as that during high load operation), and accordingly, the heat generation rate is only smaller than that during high load operation, and the description thereof is omitted here.

  FIG. 6 shows a combustion field (for example, in the combustion chamber 3 (in the case of the injector 23 having ten injection holes), which is a region in which fuel is injected during each fuel injection period when the engine 1 is operated at a low load. Specifically, it is a map (generally referred to as a φT map) showing changes in the gas temperature at each of the 10 combustion fields in the cavity 13b) and the equivalent ratio in the combustion field. That is, when the main injection (each divided main injection) is executed with the fuel injection pattern shown in FIG. 4, the combustion field of the fuel injected by the first fuel injection and the fuel injected by the second fuel injection Changes in the combustion field environment (combustion field gas temperature and equivalent ratio) in each combustion field are indicated by arrows.

  On the other hand, FIG. 7 is a φT map showing changes in the gas temperature in the combustion field, which is a region where fuel is injected during each fuel injection period, and the equivalent ratio in the combustion field during high-load operation of the engine 1. . That is, when main injection (each divided main injection) is executed with the fuel injection pattern shown in FIG. 5, the combustion field of the fuel injected by the first fuel injection and the fuel injected by the second fuel injection Changes in the combustion field environment (combustion field gas temperature and equivalent ratio) in each combustion field are indicated by arrows.

  6 and 7, when the combustion field environment reaches the smoke generation region in the figure, smoke is generated in the exhaust gas. The smoke generation region is a region where the combustion field gas temperature is relatively high and the combustion field equivalent ratio is rich. Further, when the combustion field environment reaches the NOx generation region in the figure, NOx is generated in the exhaust gas. This NOx generation region is a region where the combustion field gas temperature is relatively high and the combustion field equivalent ratio is on the lean side. Further, the X region shown in FIGS. 6 and 7 is a region where HC is likely to be generated in the exhaust gas, and the Y region is a region where CO is likely to be generated in the exhaust gas.

Hereinafter, the injection mode of each divided main injection will be described for each of the low load operation and the high load operation. The total main injection amount during low load operation is set to 18 mm ³ , for example, and the total main injection amount during high load operation is set to 40 mm ³ , for example. Incidentally, the total main injection amount during the medium load operation is set to 30 mm ³ , for example. These values are not limited to this.

(During low load operation)
As shown in FIG. 4, as each divided main injection at the time of low load operation of the engine 1, first, the timing of the first fuel injection is advanced from the compression top dead center (TDC) of the piston 13 (for example, BTDC 16 °). ), And the end of the injection is also advanced (for example, BTDC 6 °) from the compression top dead center (TDC) of the piston 13 (fuel injection control by the fuel injection control means).

  On the other hand, the timing of the second fuel injection starts injection on the advance side (for example, BTDC 4 °) from the compression top dead center (TDC) of the piston 13, and the end of injection starts from the compression top dead center (TDC) of the piston 13. Is also on the retarded side (for example, ATDC 4 °).

  A predetermined interval is provided between the first fuel injection and the second fuel injection. That is, after executing the first fuel injection, the fuel injection is temporarily stopped (the injector 23 is shut off), and the second fuel injection is started after a predetermined interval. This interval is set as, for example, the shortest valve closing period (determined by the performance of the injector 23, and the shortest period from when the injector 23 is closed until the valve opening is started: for example, 200 μs). The interval between the divided main injections is not limited to the above value, and is appropriately set so that a function in each combustion (particularly, a mutual interference cooling action described later) is exhibited as described later. Hereinafter, each fuel injection (first fuel injection and second fuel injection) will be described.

<First fuel injection>
As shown in FIG. 4, the first fuel injection at the time of the low load operation starts injection on the advance side (for example, BTDC 16 °) from the compression top dead center (TDC) of the piston 13. By starting the first fuel injection at this timing (over-advance angle timing), the volume in the combustion chamber 3 is increased in a relatively wide space (the piston 13 is at a relatively low position). ) In a situation where the in-cylinder temperature is relatively low (the in-cylinder temperature is relatively low because the degree of compression of air in the cylinder is low), fuel injection into the cylinder is performed. For this reason, in the combustion chamber 3, the fuel is in an overdiffusion state, and in the cylinder, the region where the equivalence ratio is less than 1 becomes most, and the ignition delay is promoted. During the ignition delay period, the amount of preheating in the cylinder is small.

Further, as described above, since the EGR amount is set to be relatively large (a large amount of EGR), the ratio between oxygen molecules and fuel particles in the combustion chamber 3 is low, and Since the heat in the cylinder is taken away by the CO ₂ (gas having a large heat capacity) contained in the recirculated exhaust gas, the degree of progress of the chemical reaction in the combustion chamber 3 is low. Ignition delay will be promoted.

The fuel injection amount (corresponding to the valve opening period of the injector 23) in the first fuel injection is set to 12 mm ³ , for example. This value is not limited to this. Thus, in the first fuel injection, since the fuel injection amount is set to be relatively large, the penetration force of the fuel is relatively high, and the fuel is injected over a wide range in the combustion chamber 3. In this portion, as described above, the ignition delay promoting effect by the exhaust gas (CO ₂ ) recirculated by the EGR device is sufficiently exhibited.

  In addition, there are a physical delay and a chemical delay as a fuel ignition delay in a diesel engine. The physical delay is the time required for evaporation / mixing of the fuel droplets and depends on the gas temperature of the combustion field. On the other hand, the chemical delay is the time required for chemical bonding / decomposition of fuel vapor and oxidation heat generation. As described above, in the first fuel injection, the physical delay can be greatly increased, and accordingly, the combustion hardly starts until the fuel injection timing of the second fuel injection described later.

  Further, during the injection period of the first fuel injection, a drought reduction operation for reducing the droop rate between oxygen in the cylinder and fuel spray is performed. As the ratio reduction operation, in addition to the exhaust gas recirculation operation by the EGR device described above, at least one of the intake throttle operation in the intake system and the operation of reducing the in-cylinder temperature is executed. In particular, the intake throttle operation in the intake system includes an intake throttle operation by the throttle valve 62 provided in the intake system, a supercharging reduction operation by the turbocharger 5, and an SCV (swirl not shown) provided in the intake system. Control valve) intake throttle operation and the like. Examples of the operation for decreasing the in-cylinder temperature include an operation for increasing the cooling capacity of the intercooler 61 and the EGR cooler 82, and an operation for decreasing the compression ratio in the cylinder.

  For example, when the exhaust gas recirculation operation by the EGR device is performed alone, for example, the target EGR rate is set to 30% and the opening degree of the EGR valve 81 is controlled. When the intake throttle operation is performed alone, the opening of the throttle valve 62 is reduced to, for example, 75%. Furthermore, when the operation of lowering the in-cylinder temperature, such as increasing the cooling capacity of the EGR cooler 82 and the cooling capacity of the intercooler 61, is performed, the kinetic energy of oxygen molecules and fuel particles in the cylinder will decrease. , The above-mentioned dredging rate is effectively reduced. In addition, each value mentioned above is not limited to this.

  Since the first fuel injection is executed while such a low rate reduction operation is performed, the ignition delay of the fuel injected by the first fuel injection is further promoted.

  Note that the injection amount in the first fuel injection (corresponding to the valve opening period of the injector 23) is set, for example, by experiments or simulations.

<Second fuel injection>
The second fuel injection is a final injection of a plurality of times (two in the present embodiment) main injection, and at a timing after a predetermined interval has elapsed after the first fuel injection is stopped. Executed. For example, the fuel injection is started slightly on the advance side with respect to TDC, and the fuel injection is ended on the slightly retard side with respect to TDC. Specifically, as described above, the injection is started on the advance side (for example, BTDC 4 °) from the compression top dead center (TDC) of the piston 13 and the retard side from the compression top dead center (TDC) of the piston 13. The injection has been completed (for example, ATDC 4 °). Note that the execution timing of the second fuel injection, that is, the interval between the first fuel injection and the second fuel injection, is determined by the shortest valve closing period (the performance of the injector 23 as described above, and the injector 23 is closed). However, the present invention is not limited to this, and for example, it may be adjusted according to the amount of HC generated in the combustion gas.

The fuel injection amount in the second fuel injection is set to 6 mm ³ , for example. This value is not limited to this.

  Thus, the fuel injection period of the second fuel injection is set to the compression top dead center (TDC) of the piston 13 and the period before and after that. At this timing, the volume in the combustion chamber 3 is minimized and the combustion of the fuel injected by the first fuel injection described above starts to be started, so that the temperature in the combustion chamber 3 is relatively high. The fuel ignition delay in fuel injection is substantially minimized. That is, while the ignition delay of the fuel injected by the first fuel injection is large, the ignition delay of the fuel injected by the second fuel injection is substantially minimum.

  By performing the first fuel injection and the second fuel injection in this way, the second fuel injection is performed with respect to the spray (gas lump) in the first fuel injection that has not yet started combustion or has started combustion. The spray (gas lump) in the fuel injection is united. At this time, the spray in the second fuel injection undergoes an endothermic reaction, thereby cooling the spray in the first fuel injection. Hereinafter, this cooling action is referred to as “mutual interference cooling action”.

  In this way, after the fuel injection in the second fuel injection, the temperature in the combustion chamber is kept low by the mutual interference cooling action in the combustion chamber 3, and each spray (the spray in the first fuel injection and the first fuel injection). The spraying in the two fuel injections) is combined, and the initial slow combustion, the premixed combustion, and the diffusion combustion are sequentially performed. Therefore, as shown in FIG. 4, the timing at which the heat generation rate changes from rising to falling (peak timing of the heat generation rate) exists only at one position on the retarded side (ATDC) from the compression top dead center of the piston 13. Will do.

  As a result, most of the combustion modes of the fuel injected in the first fuel injection and the second fuel injection are premixed combustion, and oxygen shortage does not occur in the combustion field, so that the generation of smoke is suppressed. Can do. Further, since the occurrence of smoke can be suppressed in this way, the EGR rate can be set even higher (for example, the EGR rate can be increased to about 60%). In other words, even if the EGR rate is set high, it is possible to obtain a situation in which smoke does not occur, and it is possible to greatly reduce the amount of NOx generated with this large amount of EGR.

  Further, when the second fuel injection is executed in a situation where the fuel injected by the first fuel injection is hardly burned (a situation without preheating), not only the fuel in the first fuel injection but also The fuel combustion mode in the second fuel injection may be premixed combustion. That is, the combustion mode in the combustion period is achieved only by the initial slow combustion and the premixed combustion. In this case, since there is no diffusion combustion, it is possible to reliably achieve both suppression of the NOx generation amount and suppression of the smoke generation amount.

  Further, by changing the ratio between the fuel injection amount of the first fuel injection and the fuel injection amount of the second fuel injection, or by changing the injection timing (degree of over-advance angle) of the first fuel injection, The magnitude of the mutual interference cooling action can also be changed. That is, it becomes possible to control the ratio of premixed combustion over the entire combustion period and the peak timing of the heat generation rate. For example, it is possible to obtain an ideal heat generation rate waveform with high efficiency in which premixed combustion is started from around TDC and the peak timing of the heat generation rate is reached at ATDC 10 °.

  The injection amount in the second fuel injection is also set by, for example, experiments or simulations.

  A change in the combustion field environment at the time of executing the main injection as described above will be described with reference to FIG. As described above, FIG. 6 is a map showing changes in the gas temperature of the combustion field and the equivalence ratio of the combustion field.

  As shown in FIG. 6, when the first fuel injection is started (point A in FIG. 6), since the fuel injection amount is relatively large, the equivalent ratio of the combustion field shifts to the rich side, while combustion occurs due to the ignition delay. Hardly proceeds, so the combustion field gas temperature rises only slightly.

  Thereafter, when the second fuel injection is started (point B in FIG. 6), the equivalence ratio of the combustion field shifts to the rich side and each combustion (initial slow combustion, premixed combustion, diffusion combustion) proceeds. As a result, the combustion field gas temperature rises. In this case, since the amount of increase in the combustion field gas temperature is limited by the mutual interference cooling action, the combustion field environment does not reach the smoke generation region or the NOx generation region.

  In the second half of this diffusion combustion, the equivalence ratio increases, but the combustion field environment is in the Y region (CO region) as shown in FIG. 6, and the generation of NOx and smoke is suppressed.

(During high load operation)
Next, the combustion mode at the time of high load operation of the engine 1 will be described.

  As shown in FIG. 5, each divided main injection during high load operation of the engine 1 includes the first fuel injection in which the slow combustion main injection and the premixed combustion main injection are continuously performed, and the diffusion combustion. And the second fuel injection, which is the main injection. In the first fuel injection (initial continuous main injection), the injection start time is set before the compression top dead center (BTDC) of the piston 13 and the injection end time is set after the compression top dead center of the piston 13 (ATDC). Has been. For example, the injection starts at 15 ° before compression top dead center (BTDC 15 °), and the injection ends at 20 ° after compression top dead center (ATDC 20 °). These values are not limited to this. Therefore, the second fuel injection is performed on the retard side with respect to the compression top dead center (TDC) of the piston 13.

Further, as described above, the fuel injection amount in the first fuel injection is set to be larger than the fuel injection amount in the second fuel injection. When the total fuel injection amount during the high load operation is 40 mm ³ , the fuel injection amount in the first fuel injection is set to 35 mm ³ and the fuel injection amount in the second fuel injection is set to 5 mm ³ , respectively. These values are not limited to this.

  A predetermined interval is provided between the first fuel injection and the second fuel injection. That is, after executing the first fuel injection, the fuel injection is temporarily stopped (the injector 23 is shut off), and the second fuel injection is started after a predetermined interval. This interval is set as, for example, the shortest valve closing period (determined by the performance of the injector 23, and the shortest period from when the injector 23 is closed until the valve opening is started: for example, 200 μs). The interval of this divided main injection is not limited to the above value, and is appropriately set so that the function in each combustion is exhibited, as will be described later. Hereinafter, each fuel injection (first fuel injection and second fuel injection) will be described.

<First fuel injection>
As shown in FIG. 5, the first fuel injection performed at the time of high load operation starts the injection on the advance side (for example, BTDC 15 °) from the compression top dead center (TDC) of the piston 13 and compresses the piston 13. The injection is terminated on the retard side (for example, BTDC 20 °) from the top dead center. That is, fuel injection at a low fuel injection rate (which has become a low fuel injection rate as the diameter of the injection hole of the injector 23 is reduced) is continued for a relatively long period of time.

  Thereby, in most of the periods other than the initial period and the end of the injection period of the first fuel injection, the fuel previously injected into the combustion chamber 3 is subsequently transferred into the combustion chamber 3 subsequently. It is cooled by the endothermic reaction of the injected fuel. By continuing such cooling action of the fuel (hereinafter referred to as “self-interference cooling action”), the combustion field temperature in the combustion chamber 3 (in the region where the spray exists) The rise in temperature is suppressed.

  Specifically, the state of the cooling action during the execution period of the first fuel injection will be described. First, in the fuel injected into the combustion chamber 3 at the initial start of the injection of the first fuel injection, the preceding fuel is not Since it does not exist, the self-interference cooling action of the fuel due to the endothermic reaction of the fuel injected at the beginning of the injection does not occur. In the fuel continuously injected after that, the endothermic reaction in the combustion chamber 3 is performed as described above, so that the fuel previously injected into the combustion chamber 3 is cooled. Become. In such a state, while the fuel injection rate of the first fuel injection is maintained substantially constant (the forward movement of the needle (movement in the valve closing direction) accompanying the stop operation of the first fuel injection), (Until the decline starts) (see the shaded period in FIG. 5). When the forward movement (movement in the valve closing direction) of the needle accompanying the stop operation of the first fuel injection is started, the penetration force of the fuel injected thereafter decreases, and the flight distance becomes shorter. As a result, the fuel spray region previously injected into the combustion chamber 3 does not reach and the self-interference cooling action does not occur. For this reason, as described above, the self-interference cooling action is performed in most of the periods excluding the initial period and the end period of the first fuel injection period.

  Further, in the first fuel injection, since the fuel injection amount is relatively large, the penetration force of the fuel is relatively high and reaches a relatively wide space in the combustion chamber 3 (the outer peripheral side space in the cavity 13b). Yes. In this portion, as described above, the combustion temperature reduction effect by the exhaust gas recirculated by the EGR device is sufficiently exhibited. Thus, since the fuel injected by the first fuel injection is cooled by the self-interference cooling action, the ignition delay is promoted, and the temperature rise in the combustion chamber 3 is suppressed.

  As a result, most of the combustion mode of the fuel injected in the first fuel injection is premixed combustion, and oxygen shortage does not occur in the combustion field, so that the generation of smoke can be suppressed. In addition, since the occurrence of smoke can be suppressed in this way, the EGR rate can be set even higher. In other words, even when the EGR rate is set high, a situation in which smoke does not occur can be obtained. Along with EGR, the amount of NOx generated can be greatly reduced.

  In addition, during the injection period of the first fuel injection, as in the case of the low-load operation described above, a reduction rate reduction operation is performed to reduce the reduction rate between the oxygen in the cylinder and the fuel spray. . That is, at least one of the exhaust gas recirculation operation by the EGR device, the intake throttle operation in the intake system, and the operation of lowering the in-cylinder temperature is performed. Since these percentage reduction operations have been described above, a description thereof is omitted here.

  Note that the ratio reduction operation is not limited to the above-described operation, and may include an operation of retarding the fuel injection timing from the injector 23. In this case, each waveform shown in FIG. 5 shifts to the retard side, and the first fuel injection period and the second fuel injection period also shift to the retard side. Thus, the combustion temperature in the combustion chamber 3 can be lowered by shifting the fuel injection timing to the retard side.

  Thus, since the first fuel injection is performed while the reduction rate reduction operation is performed, the temperature rise in the combustion chamber 3 is suppressed by the synergistic effect with the self-interference cooling action, and the generation of smoke and the large amount of smoke are suppressed. The amount of NOx generated due to EGR can be reduced.

  The injection amount (corresponding to the valve opening period of the injector 23) and the injection period in the first fuel injection and the injection period are set by, for example, experiments or simulations.

<Second fuel injection>
The second fuel injection during high-load operation of the engine 1 is performed at a timing when the temperature in the cylinder exceeds the temperature (900K) at which diffusion combustion is possible after the fuel injected by the first fuel injection is premixed and combusted. Is started. For example, the injection is performed at a timing of ATDC 22 °. And the fuel injected by this 2nd fuel injection becomes diffusion combustion which burns sequentially immediately after the injection.

The fuel injection amount in the second fuel injection is set to 5 mm ³ , for example, as described above. This value is not limited to this. Note that the fuel injection amount in the second fuel injection is set, for example, so that the combustion noise during the diffusion combustion is equal to or lower than a preset allowable combustion noise. That is, the remaining fuel amount obtained by subtracting the fuel injection amount of the second fuel injection set as described above from the total fuel injection amount during high load operation (40 mm ^{3 in the} present embodiment) is The fuel is injected into the combustion chamber 3 by the first fuel injection.

  The fuel injected by the second fuel injection has a lower penetrating force than the fuel injected by the first fuel injection, and therefore does not reach the combustion field of the fuel injected by the first fuel injection. In a relatively narrow area (inner circumferential space in the cavity 13b), diffusion combustion is performed in a combustion field independent of the combustion field of the first fuel injection.

  In the above description, the case where the combustion field of the first fuel injection and the combustion field of the second fuel injection are independent from each other has been described. However, there is a case where parts of these regions overlap each other.

  Further, since the combustion in the second fuel injection is diffusion combustion, if the fuel injection timing is controlled, the heat generation rate waveform accompanying the combustion can be controlled.

As described above, the fuel injection amount in the second fuel injection for performing the diffusion combustion is set to be significantly smaller than the fuel injection amount in the first fuel injection. In other words, most of the total fuel injection amount (for example, 40 mm ³ ) that is set relatively large during the high load operation is the fuel injection amount in the first fuel injection, and the fuel injection amount in the second fuel injection is greatly increased. Set it less. For this reason, the ratio of the combustion of fuel injected by the first fuel injection (mostly premixed combustion) is ensured, and the ratio of the combustion of fuel injected by the second fuel injection (diffusion combustion) is significantly reduced. It will be set. By setting the amount of fuel that performs diffusion combustion in such a small amount, it is possible to suppress the generation of smoke and reduce the amount of NOx generated.

  The injection amount in the second fuel injection is also set by, for example, experiments or simulations.

  Changes in the combustion field environment during main injection execution during high load operation as described above will be described with reference to FIG. As described above, FIG. 7 is a map showing changes in the gas temperature of the combustion field and the equivalence ratio of the combustion field.

  As shown in FIG. 7, when the first fuel injection is started (point A ′ in FIG. 7), the fuel injection amount is relatively large, so the equivalence ratio of the combustion field increases. At this time, initial slow combustion and premixed combustion are performed in the combustion field. However, since the self-interference cooling action occurs in the combustion field as described above, the combustion field gas temperature rises only slightly. Therefore, this combustion field environment does not reach the smoke generation region or the NOx generation region.

  Thereafter, when the second fuel injection is started (point B ′ in FIG. 7), diffusion combustion is started in the cylinder and the combustion field gas temperature is increased. In the diffusion combustion in this case, the combustion field environment does not reach the smoke generation region or the NOx generation region.

  In the second half of this diffusion combustion, the combustion field environment is the Y region (CO region) as shown in FIG. 7, and the generation of NOx and smoke is suppressed.

-Fuel injection mode change control-
Next, the control operation for changing the fuel injection mode in accordance with the load state of the engine 1 as described above will be described with reference to the flowchart of FIG. The routine shown in FIG. 8 is executed every predetermined time after the engine is started, for example, every several milliseconds.

  First, in step ST1, an engine load determination operation is performed. For example, in the ROM 102, a load determination map and an arithmetic expression using the engine speed calculated based on the detection value of the crank position sensor 40, the accelerator pedal depression amount obtained from the detection signal of the accelerator opening sensor 47, and the like as parameters are stored in the ROM 102. The engine load is judged by memorizing it.

  Then, it moves to step ST2 and determines whether an engine load is a low load state. If the engine load is in a high load state and NO is determined in step ST2, the process proceeds to step ST3 to execute divided main injection (spray self-interference cooling injection) for performing the above-described self-interference cooling action. That is, after the first fuel injection shown in FIG. 5 in which the slow combustion main injection and the premixed combustion main injection are continuously performed is executed near the compression top dead center (TDC), the fuel injection is temporarily performed. The divided main injection is executed such that the second fuel injection is started after stopping and passing through a predetermined interval.

  On the other hand, when the engine load is in a low load state and YES is determined in step ST2, the process proceeds to step ST4 to perform divided main injection (spray mutual interference cooling injection) for performing the above-described mutual interference cooling action. Execute. That is, after the first fuel injection shown in FIG. 4 is performed on the advance side (BTDC) from the compression top dead center, the fuel injection is temporarily stopped, and after a predetermined interval, the second fuel injection is performed. Divided main injection is executed such that it is executed near the compression top dead center (TDC).

  In this way, the fuel injection mode is changed according to the load state.

  As described above, according to the present embodiment, when the engine 1 is operated at a low load, the mutual interference cooling action performed between the fuel injected by the first fuel injection and the fuel injected by the second fuel injection. Thus, the ignition delay in the combustion chamber 3 is promoted, thereby suppressing the temperature rise in the combustion chamber 3. For this reason, a large proportion of premixed combustion during the combustion period can be secured, and the proportion of diffusion combustion is set to be significantly small. Even if a large amount of EGR is performed (for example, even if the EGR rate is set to 60%), smoke Since generation | occurrence | production of this can be suppressed, coexistence with suppression of NOx generation amount and suppression of smoke generation amount can be aimed at. Further, since the NOx generation amount can be suppressed without retarding the fuel injection timing, it is possible to improve the combustion efficiency by setting the fuel injection timing to the advance side.

  Further, during a high load operation, a relatively large part of the total fuel injection amount set is set as the fuel injection amount in the first fuel injection, and the fuel injection amount in the second fuel injection is set to be significantly small. . Further, the ignition delay in the combustion chamber 3 is promoted by the self-interference cooling action of the fuel injected in the first fuel injection, thereby suppressing the temperature rise in the combustion chamber 3. Also in this case, a large proportion of premixed combustion during the combustion period is ensured, and the proportion of diffusion combustion, which is the combustion of fuel injected in the second fuel injection, is set to be significantly small. In addition, since the generation of smoke can be suppressed, it is possible to achieve both the suppression of the NOx generation amount and the suppression of the smoke generation amount. Further, since the NOx generation amount can be suppressed without retarding the fuel injection timing, the combustion efficiency can be improved by setting the fuel injection timing to the advance side.

  Further, in the present embodiment, the NOx generation amount can be greatly reduced, so that the NSR catalyst 75 and the DPNR catalyst 76 can be downsized, and the NOx generation amount can be made substantially “0”. Therefore, it is possible to eliminate the NSR catalyst 75 and the DPNR catalyst 76 and replace the three-way catalyst in the exhaust system 6 instead. According to this, it is possible to realize an exhaust system having a relatively simple configuration similar to that of a gasoline engine in a diesel engine.

-Modification-
Next, a modified example of the present invention will be described. In this modification, by executing three divided main injections as the main injection mode, the total main injection amount (total fuel injection amount for obtaining the required torque) required for the main injection is secured. It is what you do. That is, the first fuel injection is divided into two fuel injections, and then the second fuel injection is executed. In the following description, the advanced angle side fuel injection of the two divided first fuel injections is referred to as “advanced angle side first fuel injection”, and the retarded angle side fuel injection is referred to as “retarded angle side first fuel injection”. I will call it.

  FIG. 9 shows the change in the heat generation rate in the cylinder and the fuel injection pattern during the low-load operation of the engine 1 and during the execution period of the main injection in the present modification.

  As shown in FIG. 9, as each divided main injection at the time of low load operation of the engine 1, the compression top dead center of the piston 13 for each of the advance side first fuel injection and the retard side first fuel injection. Injection is started on the advance side with respect to (TDC), and the end of injection is also set on the advance side with respect to the compression top dead center (TDC) of the piston 13. For example, the advance side first fuel injection is started at BTDC 18 ° on the advance side of the compression top dead center of the piston 13 and terminated at BTDC 14 °. Further, the retarded side first fuel injection is started at BTDC 12 ° which is an advance side of the compression top dead center of the piston 13 and is terminated at BTDC 6 °.

  On the other hand, as the injection timing of the second fuel injection, the injection is started on the advance side (for example, BTDC 4 °) from the compression top dead center (TDC) of the piston 13, and the injection end is the compression top dead center (TDC) of the piston 13. ) On the more retarded side (for example, ATDC 4 °).

  Further, the total fuel injection amount injected by the advance side first fuel injection and the retard side first fuel injection is set to be larger than the fuel injection amount injected by the second fuel injection.

  Even when such three divided main injections are executed, the fuel injected by the advance side first fuel injection and the retard side first fuel injection is injected by the first fuel injection in the above-described embodiment. As in the case of the fuel, the ignition delay is accelerated, and the ignition delay in the combustion chamber 3 is promoted by the mutual interference cooling action performed with the fuel injected by the second fuel injection. . As a result, the temperature rise in the combustion chamber 3 is suppressed, the ratio of premixed combustion during the combustion period can be ensured large, and the ratio of diffusion combustion is set to be significantly small. Since generation | occurrence | production can be suppressed, coexistence with suppression of NOx generation amount and suppression of smoke generation amount can be aimed at.

  Further, by increasing the number of divisions of main injection in this way, the peak value of the heat generation rate can be suppressed low, and combustion noise can be suppressed.

-Other embodiments-
In the embodiment and the modification described above, the case where the present invention is applied to an in-line four-cylinder diesel engine mounted on an automobile has been described. The present invention is applicable not only to automobiles but also to engines used for other purposes. Further, the number of cylinders and the engine type (separate type engine, V-type engine, horizontally opposed engine, etc.) are not particularly limited.

  In the above-described embodiment and modification, the NSR catalyst 75 and the DPNR catalyst 76 are provided as the manipulator 77. However, the NSR catalyst 75 and a DPF (Diesel Particle Filter) may be provided.

  In the embodiment and the modification, the EGR device is configured to recirculate the exhaust gas in the exhaust manifold 72 to the intake system 6. The present invention is not limited to this, and an LPL (Low Pressure Loop) EGR device that recirculates exhaust gas downstream of the turbine wheel 52 in the turbocharger 5 to the intake system 6 may be employed.

  In the above embodiment and the modification, the injection timing of the second fuel injection performed during the low load operation is set near TDC (compression top dead center of the piston 13). The present invention is not limited to this, and the injection timing of the second fuel injection may be set to ATDC (a retarded side from the compression top dead center of the piston 13).

  Moreover, the embodiment and the modification described above have described the engine 1 to which the piezo injector 23 that changes the fuel injection rate by being in a fully opened valve state only during the energization period. The present invention is not limited to this, and can also be applied to an engine equipped with a variable injection rate injector capable of changing the injection rate according to the energization voltage value or the like.

  Moreover, the embodiment and the modification described above have described the engine 1 to which the injector 23 having a minute nozzle hole diameter is applied. The present invention is not limited to this, and can also be applied to an engine equipped with an injector having a general nozzle hole diameter.

  Further, in the above-described embodiment and modification, the engine 1 has been described that has a low fuel injection rate as the nozzle hole diameter of the injector 23 is reduced. The present invention is not limited to this, and can also be applied to an engine having a low fuel injection rate because the fuel injection pressure (common rail internal pressure) is set low.

  Further, in the embodiment and the modification described above, after injection and post injection are not executed as sub injection after execution of the second fuel injection. The present invention is not limited to this, and at least one of these after injection and post injection may be executed.

  In addition, although the injector 23 in the above-described embodiment and the modification includes 10 injection holes, the number of injection holes can be arbitrarily set. However, it is necessary to set the number of injection holes so that a necessary fuel injection amount can be secured within a predetermined fuel injection period. In particular, in an injector having a small nozzle hole diameter, it is necessary to set a large number (for example, 10 or more) of nozzle holes.

  In the above embodiment, the fuel injection amount in the first fuel injection is set to be larger than the fuel injection amount in the second fuel injection. The present invention is not limited to this, and in the above embodiment, the fuel injection amount in the first fuel injection and the fuel injection amount in the second fuel injection are set substantially the same, or the fuel in the second fuel injection is set. A larger amount of injection may be set.

  The present invention can be applied to fuel injection control in a common rail in-cylinder direct injection multi-cylinder diesel engine mounted on an automobile.

1 engine (internal combustion engine)
13 Piston 3 Combustion chamber 23 Injector (fuel injection valve)
6 Intake system 62 Throttle valve (intake throttle valve)
7 Exhaust system 8 EGR passage 81 EGR valve 82 EGR cooler

Claims (11)

An exhaust gas recirculation device that recirculates part of the exhaust gas discharged to the exhaust system to the intake system and at least initial slow combustion and premixed combustion are sequentially performed as main injection, which is fuel injection for generating torque. In a combustion control device for a compression ignition type internal combustion engine provided with a fuel injection valve configured to be able to execute a plurality of divided main injections including fuel injection for
With the exhaust gas recirculation device, a part of the exhaust gas is recirculated to the intake system and the pre-heating operation before the start of the initial slow combustion is not performed. The fuel injection is performed near the compression top dead center, and the fuel injection performed before the final divided main injection is performed on the advance side from the compression top dead center of the piston. The gas mass in the fuel injection and the gas mass in the final split main injection are merged, and the timing at which the heat generation rate changes from rising to falling seems to exist only at one position on the retard side of the compression top dead center. A combustion control apparatus for an internal combustion engine, further comprising fuel injection control means for executing fuel injection control. An exhaust gas recirculation device that recirculates part of the exhaust gas discharged to the exhaust system to the intake system and at least initial slow combustion and premixed combustion are sequentially performed as main injection, which is fuel injection for generating torque. In a combustion control device for a compression ignition type internal combustion engine provided with a fuel injection valve configured to be able to execute a plurality of divided main injections including fuel injection for
With the exhaust gas recirculation device, a part of the exhaust gas is recirculated to the intake system and the pre-heating operation before the start of the initial slow combustion is not performed. The fuel injection is performed near the compression top dead center, and the fuel injection performed before the final divided main injection is performed on the advance side from the compression top dead center of the piston. A combustion control device for an internal combustion engine, comprising fuel injection control means for performing fuel injection control for causing an interference cooling action by the gas mass in the final divided main injection with respect to the gas mass in the fuel injection. In the internal combustion engine combustion control apparatus according to claim 1 or 2,
The combustion control device for an internal combustion engine, wherein the fuel injection control means is configured to set a period from before the compression top dead center to after the compression top dead center as a period of the final divided main injection. . In the internal combustion engine combustion control apparatus according to claim 1 or 2,
The fuel injection control means causes the fuel injection, which is performed at a stage prior to the final divided main injection, to be advanced from the compression top dead center of the piston during low-load operation of the internal combustion engine. Is executed in the vicinity of the compression top dead center of the piston, while at the time of high load operation of the internal combustion engine, fuel injection performed in a stage before the final divided main injection is executed in the vicinity of the compression top dead center of the piston. A combustion control device for an internal combustion engine, characterized in that the main injection is executed on the retard side from the compression top dead center of the piston. In the internal combustion engine combustion control apparatus according to claim 1 or 2,
The fuel injection control means sets the penetration force of the fuel injected in the final divided main injection to be lower than the penetration force of the fuel injected in the fuel injection performed in a stage before the final divided main injection. A combustion control apparatus for an internal combustion engine, characterized in that it is configured. In the internal combustion engine combustion control apparatus according to claim 5,
The fuel injection control means sets the total fuel injection amount injected in the final divided main injection to be smaller than the total fuel injection amount injected in the fuel injection performed in a stage before the final divided main injection. A combustion control apparatus for an internal combustion engine, characterized in that it is configured. In the internal combustion engine combustion control apparatus according to claim 1 or 2,
The fuel injection control means is an injection timing of fuel injection performed before the final divided main injection so that substantially the entire fuel injected by the final divided main injection becomes initial slow combustion and premixed combustion. And a combustion control apparatus for an internal combustion engine, wherein the combustion control apparatus is configured to set an injection amount. In the internal combustion engine combustion control apparatus according to claim 1 or 2,
An internal combustion engine that is configured to perform fuel injection that is performed before the final divided main injection while performing a rate reduction operation that reduces the rate of oxygen and fuel spray in the cylinder. Engine combustion control device. In the internal combustion engine combustion control apparatus according to claim 7,
Combustion of an internal combustion engine characterized in that at least one of exhaust gas recirculation operation by the exhaust gas recirculation device, intake throttle operation in the intake system, and operation of lowering the in-cylinder temperature is executed as the reduction ratio operation Control device. In the internal combustion engine combustion control apparatus according to claim 1 or 2,
A combustion control device for an internal combustion engine, characterized in that an exhaust gas recirculation rate in a state where a part of exhaust gas is recirculated to the intake system by the exhaust gas recirculation device is set to 20% or more. In the internal combustion engine combustion control apparatus according to claim 1 or 2,
A combustion control apparatus for an internal combustion engine, wherein the fuel injection valve has an injection hole diameter reduced so that a fuel injection rate when the valve is opened is equal to or less than a predetermined injection rate.

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