JP3134766B2 - In-cylinder internal combustion engine - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

The present invention relates to a direct injection internal combustion engine.

[0002]

2. Description of the Related Art Fuel is injected into a combustion chamber during a compression stroke or an intake stroke approximately 60 degrees before a compression top dead center, and the average particle diameter of the injected fuel at this time is determined by the fact that the temperature of the fuel particles rises substantially above the compression stroke. After the dead center or the compression top dead center, the particle size must reach the boiling point of the main fuel component determined by the pressure at that time,
A compression ignition type that prevents evaporation of fuel due to boiling from the fuel particles until it almost reaches the compression top dead center after injection, and makes the fuel of the fuel particles boil and evaporate after the compression top dead center to ignite and burn the fuel. BACKGROUND ART Internal combustion engines are known (see Japanese Patent Application Laid-Open No. 7-317588). In this compression ignition type internal combustion engine, the generation amounts of soot and NOx become almost zero.

[0003]

However, even with this compression ignition type internal combustion engine, if the temperature of the intake gas in the combustion chamber becomes too high at the end of the compression stroke, soot and NO
If x occurs and the temperature of the intake gas in the combustion chamber at the end of the compression stroke becomes too low, good ignition is not performed, and a problem arises that unburned HC is generated.

[0004]

According to a first aspect of the present invention, fuel is injected into a combustion chamber during a compression stroke or an intake stroke substantially before 60 degrees before a compression top dead center. The average particle size of the injected fuel at this time should be equal to or larger than the particle size at which the temperature of the fuel particles reaches the boiling point of the main fuel component determined by the pressure at the compression top dead center or after the compression top dead center. In a cylinder-injection internal combustion engine in which the evaporation of fuel due to boiling from the fuel particles is prevented until reaching the point, and the fuel of the fuel particles is boiled and evaporated to burn and burn the fuel substantially after the compression top dead center.
Temperature reduction means for lowering the temperature of the intake gas sucked into the combustion chamber at the end of the compression stroke in order to start combustion after 10 degrees before the compression top dead center, and when the engine load is lower than a predetermined load. When the engine load is too high, the temperature of the intake gas at the end of the compression stroke is lowered by the temperature lowering means, and when the engine load is lower than a predetermined load, the action of lowering the temperature of the intake gas by the temperature lowering means is stopped. That is, when the engine load becomes high, the temperature of the intake gas in the combustion chamber at the end of the compression stroke becomes too high, and combustion is started after 10 degrees before the compression top dead center.
In this case, after 10 degrees before top dead center
The temperature of the intake gas in the combustion chamber at the end of the compression stroke is lowered by the temperature lowering means so that the combustion starts .

In the second invention, in the first invention,
Temperature increasing means for increasing the temperature of the sucked intake gas at the end of the compression stroke; and when the engine load is lower than a predetermined set value, the temperature of the intake gas is increased by the temperature increase at the end of the compression stroke. When the engine load is higher than a predetermined set value, the operation of increasing the temperature of the intake gas by the temperature increasing means is stopped. In other words, when the engine load is low, the temperature of the intake gas in the combustion chamber at the end of the compression stroke becomes too low, so that good ignition cannot be obtained. The temperature of the intake gas in the combustion chamber is increased.

[0006]

1 to 3 show a case where the present invention is applied to a four-stroke compression ignition internal combustion engine. However, the present invention can also be applied to a spark ignition type gasoline engine. Referring to FIGS. 1 to 3, 1 is an engine body, 2 is a cylinder block, 3 is a cylinder head,
Reference numeral 4 denotes a piston, 5 denotes a combustion chamber, 6 denotes a pair of intake valves, 7 denotes a pair of intake ports, 8 denotes a pair of exhaust valves, 9 denotes a pair of exhaust ports, and 10 denotes a top central portion of the combustion chamber 5. Each shows a fuel injection valve. Each intake port 7 has an intake manifold 11
Are connected to the air cleaner 12 through the exhaust ports 9.
Is connected to an exhaust pipe 14 via an exhaust manifold 13. The exhaust manifold 13 and the collecting portion of the intake manifold 11 are connected to each other by an exhaust gas recirculation (hereinafter, referred to as EGR) passage 15, and an EGR passage 15 is provided in the EGR passage 15.
A control valve 16 is arranged. The EGR control valve 16 is controlled based on an output signal of the electronic control unit 30.

As shown in FIG. 1, a cooling device 17 for cooling EGR gas is disposed in the EGR passage 15. The cooling device 17 has a cooling water inlet 18 and a cooling water outlet 1
The cooling water outlet 19 is connected to a cooling water outflow passage 22 of a radiator 21 via a return conduit 20.
On the other hand, the cooling water outlet passage 22 of the radiator 21 is connected via a supply conduit 23 to an electromagnetic switching valve 24 which is connected on the one hand to a return conduit 20 via a bypass conduit 25 which bypasses the cooling device 17. On the other hand, it is connected to a cooling water inlet 18. When the supply conduit 23 is connected to the cooling water inlet 18 by the switching operation of the electromagnetic switching valve 24, the cooling water is supplied to the cooling device 17, and thus the cooling operation of the EGR gas is performed. On the other hand, when the supply conduit 23 is connected to the bypass conduit 25 by the switching operation of the electromagnetic switching valve 24, the cooling operation of the EGR gas is stopped.

On the other hand, as shown in FIGS. 2 and 3, each intake port 7 comprises a straight port extending substantially straight. Therefore, in the compression ignition type internal combustion engine shown in FIGS. No swirl is generated in the combustion chamber 5 by the airflow flowing into the combustion chamber.
As shown in FIG. 1, the electronic control unit 30 is composed of a digital computer, and a ROM (Read Only Memory) 32, R
It has an AM (random access memory) 33, a CPU (microprocessor) 34, an input port 35 and an output port 36. An air-fuel ratio sensor 26 for detecting the air-fuel ratio in the combustion chamber 5 from the oxygen concentration in the exhaust gas is disposed in the exhaust manifold 13, and an output signal of the air-fuel ratio sensor 26 is input via a corresponding AD converter 37. Input to port 35. The accelerator pedal 27 is connected to a load sensor 28 that generates an output voltage proportional to the amount of depression of the accelerator pedal 27. The output voltage of the load sensor 28 is input to an input port 35 via a corresponding AD converter 37. You. Further, the input port 35 has a crank angle sensor 2 for generating an output pulse every time the engine rotates at a constant crank angle.
9, the current crank angle and the engine speed are calculated from the output pulse of the crank angle sensor 29. On the other hand, the output port 36 is connected to each fuel injection valve 10, the EGR control valve 16 and the electromagnetic switching valve 2 via the corresponding drive circuit 38.
4 is connected.

FIG. 4 is a side sectional view of the fuel injection valve 10. Referring to FIG. 4, reference numeral 41 denotes a needle slidable in the injection valve body 40, reference numeral 42 denotes a valve body integrally formed at the tip of the needle 41, reference numeral 43 denotes a spring retainer attached to the upper end of the needle 41, reference numeral 44 Is a compression spring that urges the spring retainer 43 upward, 45 is a rod aligned with the needle 41, 46 is a fuel storage chamber having a volume several tens of times the fuel injection amount at the maximum injection amount, and 47 is a fuel storage chamber. A piston slidably disposed within the injection valve body 40, a piezoelectric element 48 for driving the piston 47, a disc spring 49 for urging the piston 47 toward the piezoelectric element 48, and a piston 47 for the piston 47. The variable volume chambers 51 defined by the top surfaces of the fuel supply ports 51 respectively indicate fuel supply ports.

A fuel supply port 51 is provided with an injection pump (not shown).
The fuel discharged from the injection pump is supplied to the fuel supply port 51. The fuel supplied to the fuel supply port 51 is supplied to the fuel storage chamber 46 via a check valve 52 that can only flow from the fuel supply port 51 to the fuel storage chamber 46. On the one hand, the fuel supplied into the fuel storage chamber 46 is guided around the distal end of the needle 41 via the fuel passage 53, and on the other hand, a check which can flow only from the fuel storage chamber 46 to the variable volume chamber 50 is provided. It is supplied to the variable volume chamber 50 via the valve 54. The fuel supplied into the variable volume chamber 50 is guided to the top surface 55 of the rod 45, so that the fuel pressure in the variable volume chamber 50 acts on the top surface 55 of the rod 45.

FIG. 5 shows the tip of the needle 41. As shown in FIG. 5, the valve body 42 of the needle 41 has a conical injection fuel guide surface 42a, and the injection fuel guide surface 42a is normally placed on the seat surface 56 by the spring force of the compression spring 44 (FIG. 4). I'm sitting. At this time, the fuel injection valve 10
Is stopped. Fuel injection valve 1
When fuel is to be injected from 0, the piezoelectric element 48 is charged. When electric charge is charged in the piezoelectric element 48, the piston 47 is lowered because the piezoelectric element 48 extends in the axial direction. Piston 47
Is lowered, the fuel pressure in the variable volume chamber 50 increases, and the rod 45 is pushed down, so that the needle 41 descends and the valve body 42 separates from the seat surface 56.
As a result, the fuel in the fuel storage chamber 46 is injected from between the valve body 42 and the seat surface 56.

Next, when the electric charge charged in the piezoelectric element 48 is discharged, the piezoelectric element 4
8 contracts in the axial direction, and the piston 47 rises. As a result, since the fuel pressure in the variable volume chamber 50 decreases, the rod 45 and the needle 41 rise due to the spring force of the compression spring 44, and the injection fuel guide surface 42a of the valve body 42 is seated on the seat surface 56 again. . In this manner, the fuel injection operation is stopped.

As shown in FIG. 5, during fuel injection, the injected fuel F is guided by the injected fuel guide surface 42a of the valve body 42 and spreads out from the tip of the needle 41, that is, from the nozzle opening of the fuel injection valve 10 into a thin film cone. Can be In the embodiment shown in FIGS. 1 to 3, the fuel injection valve 10 is disposed at the center of the top of the combustion chamber 5. Therefore, in this embodiment, the fuel F is supplied to the center of the top of the combustion chamber 5 as shown in FIG. From the combustion chamber 5
Is sprayed so as to spread in the shape of a thin film cone toward the peripheral portion.

Next, the basic combustion method of the present invention will be described with reference to FIGS. Note that this combustion method will be described focusing on high load operation in which soot and NOx are most likely to be generated. As long as the fuel is atomized and injected so that the average particle diameter of the fuel particles becomes 50 μm or less as in the conventional case, no matter how the injection timing is set or the fuel injection pressure is set, N
It is difficult to reduce Ox at the same time, and it is impossible to make the generation amount of ease and NOx substantially zero. This is because the conventional combustion method has an essential problem. That is, in the conventional combustion method, as soon as the fuel is injected due to the small particle size of the fuel particles, a part of the fuel is vaporized immediately, and the vaporized fuel starts rapid combustion at an early stage. As described above, when rapid combustion is started early after the start of injection, the following injected fuel jumps into the combustion flame, so that these injected fuels are burned in a state of insufficient air, so that soot Will happen. Also, when the injected fuel burns rapidly and rapidly and the combustion pressure rises sharply, the combustion temperature rises and thus NO
x will occur.

However, the average particle size of the injected fuel is made much larger than the average particle size used in the conventional combustion method, and the injection timing is considerably longer than the injection timing normally used in the conventional combustion method. Soot and N
It has been found that the generation amount of Ox can be reduced to substantially zero. Next, this will be described. The curve in FIG. 6 shows a change in the pressure P in the combustion chamber 5 due to only the compression action of the piston 4. As can be seen from FIG.
The pressure P inside rises rapidly when the pressure exceeds BTDC 60 degrees before the top dead center of compression. This is independent of the opening timing of the intake valve 6, and the pressure P in the combustion chamber 5 changes as shown in FIG. 6 in any reciprocating internal combustion engine.

The curve shown by the solid line in FIG. 7 shows the boiling temperature of the fuel at each crank angle, that is, the boiling point T. If the pressure P in the combustion chamber 5 rises, the boiling point T of the fuel also rises with it.
When it exceeds C60 degrees, it rises rapidly. On the other hand, the broken line in FIG. 7 indicates the difference in the temperature change of the fuel particles due to the difference in the diameter of the fuel particles when the fuel is injected at BTDC θ ₀ degrees before the compression top dead center. The temperature of the fuel particles immediately after the injection is lower than the boiling point T determined by the pressure at that time, and then the fuel particles receive heat from the surroundings and rise in temperature. At this time, the temperature rise speed of the fuel particles increases as the particle size decreases.

That is, the particle size of the fuel particles is from 20 μm to 50 μm.
If it is about μm, the temperature of the fuel particles rapidly rises after the injection and reaches the boiling point T at a crank angle far before the compression top dead center TDC, and a rapid vaporization of fuel due to boiling from the fuel particles starts. Is done. Also, as can be seen from FIG. 7, even when the particle size of the fuel particles is 200 μm, the temperature of the fuel particles reaches the boiling point T before reaching the compression top dead center TDC, and the rapid vaporization of fuel due to boiling starts. If the rapid fuel evaporation action due to boiling is started before reaching the compression top dead center TDC, explosive combustion occurs at this time due to the evaporated fuel, and thus a large amount of soot and NOx as described above. Will occur.

On the other hand, when the diameter of the fuel particle is larger than about 500 μm, the temperature rise rate of the fuel particle is slowed down. Do not reach. Therefore, if the diameter of the fuel particles is made larger than about 500 μm, the fuel will not evaporate rapidly due to boiling before it reaches the compression top dead center TDC, but will be substantially boiling at the compression top dead center TDC or after the compression top dead center TDC. A rapid fuel vaporization action is started. Therefore, when the diameter of the fuel particles is larger than about 500 μm, the generation of soot and NOx based on the boiling evaporation of the fuel before the compression top dead center TDC can be prevented.

Incidentally, the fuel actually contains various components having different boiling points, and the boiling point of the fuel means that there are many boiling points. Therefore, when considering the boiling point of the fuel, it is preferable to consider the boiling point of the main fuel component.
In addition, since the particle size of the injected fuel cannot be completely uniform, it is preferable to consider the average particle size of the injected fuel when considering the particle size of the injected fuel. Considering this, the average particle diameter of the injected fuel should be larger than the particle diameter at which the temperature of the fuel particles reaches the boiling point T of the main fuel component, which is determined by the pressure at the compression top dead center TDC or after the compression top dead center TDC. For example, the fuel does not rapidly evaporate due to boiling from the fuel particles until it almost reaches the compression top dead center TDC after the injection, and abrupt evaporation due to the boiling from the fuel particles occurs almost after the compression top dead center TDC.

However, in order to make the generation amount of soot and NOx almost zero, the diameter of the fuel particles is made larger than about 500 μm, and the injected fuel particles are evenly dispersed in the combustion chamber 5. It is necessary. Next, this will be described with reference to FIG. In FIG. 8, X indicates fuel particles. As described above, the fuel is injected early in the compression stroke, and the fuel particles X
When the diameter is larger than about 500 μm, the vaporization of the fuel by the boiling from the fuel particles X is prevented until the compression top dead center TDC is reached. However, even before the compression top dead center TDC is reached, the fuel evaporating action is not performed due to boiling. Therefore, when the fuel is injected, a layer of the evaporated fuel is formed around the fuel particles X. become.

On the other hand, as the compression stroke proceeds, the temperature in the combustion chamber 5 rises, and when the temperature in the combustion chamber 5 exceeds a certain temperature, the fuel vapor around the fuel particles X is combined with oxygen and burned. That is, combustion by the oxidation reaction of the evaporated fuel around the fuel particles X is started. At this time, a region to which the combustion heat of the evaporated fuel reaches around each fuel particle X is indicated by a broken line Y around each fuel particle X in FIG. When the dispersion density of the fuel particles X is low as shown in FIG. 8A, the regions Y are separated from each other, and when the dispersion density of the fuel particles X is high as shown in FIG. overlap.

As shown in FIG. 8B, when the regions Y overlap each other, the temperature of the space region between the fuel particles X increases due to the heat of combustion of the evaporated fuel around the fuel particles X. As described above, when the temperature of the space region between the fuel particles X increases, the temperature of the fuel particles X increases, and as a result, hydrocarbons in the fuel particles X are thermally decomposed into hydrogen molecules H ₂ , carbon C, and methane CH ₄ . As a result, when the temperature in the combustion chamber 5 rises, the hydrogen molecules H
₂ burns explosively, and the temperature in the combustion chamber 5 becomes extremely high, so that a large amount of NOx is generated. Further, when carbon C is generated by thermal decomposition, these carbons are bonded to each other to generate soot.

On the other hand, when the regions Y are separated from each other as shown in FIG. 8A, the heat of combustion of the evaporated fuel around the fuel particles X is not transmitted to the regions Y around the other fuel particles X.
As a result, the temperature of each fuel particle X does not increase so much.
As a result, the hydrocarbons in the fuel particles X do not undergo thermal decomposition, so that explosive combustion by the hydrogen molecules H ₂ does not occur, and the temperature in the combustion chamber 5 does not become extremely high.
Ox generation will be prevented. Further, since carbon C is not generated by the thermal decomposition, the carbon is not bonded to each other and does not grow until soot. As shown in FIG. 8A, if the fuel particles X are uniformly dispersed and the dispersion density of the fuel particles X is reduced, the generation of soot and NOx can be prevented.

As shown in FIG. 8A, in order to uniformly disperse the fuel particles X in the entire combustion chamber 5, fuel is injected from the fuel injection valve 10 when the pressure P in the combustion chamber 5 is low. I have to do it. That is, when the pressure P in the combustion chamber 5 increases, the air resistance increases, and the flight distance of the injected fuel becomes shorter.
As shown in (A), the fuel particles cannot spread all over the combustion chamber 5. As described above, when the pressure P in the combustion chamber 5 exceeds BTDC 60 degrees before the compression top dead center, the pressure P rapidly rises and becomes high, and in fact, the BTDC 60 before the compression top dead center DTDC 60 increases.
If the fuel injection is performed after exceeding the degree, the fuel particles do not sufficiently spread into the combustion chamber 5 as shown in FIG. On the other hand, the pressure P in the combustion chamber 5 is low before the BTDC approximately 60 degrees before the compression top dead center, so that the BTDC before the compression top dead center is almost zero.
If the fuel injection is performed before 60 degrees, the fuel particles are uniformly dispersed throughout the combustion chamber 5 as shown in FIG. 9B. Therefore, by performing the fuel injection before the BTDC 60 degrees before the compression top dead center, the generation of soot and NOx is prevented. In this case, the fuel injection timing may be during the compression stroke or the intake stroke as long as the BTDC is before 60 degrees before the compression top dead center.

An important point in carrying out this combustion method is to disperse the fuel having a large particle diameter throughout the combustion chamber 5 while keeping a space between the fuel particles, and therefore, from the viewpoint of hardware. The fuel injection valve 10 plays an important role in executing this combustion method. FIG. 4 shows an example of a fuel injection valve 10 suitable for executing this combustion method. In this fuel injection valve 10, the fuel injection pressure is set to a low pressure of about 20 MPa so as to increase the diameter of the fuel particles. Have been.

As described above, the diameter of the fuel particles is set to be larger than about 500 μm and the fuel injection timing is set substantially before the compression top dead center B
If the TDC is 60 degrees or less, the fuel vaporization due to boiling starts substantially after the compression top dead center TDC or after the compression top dead center TDC, and as shown in FIG.
And the amount of NOx generated can be reduced to almost zero.

As described above, when the fuel is injected at an early stage of the compression stroke and the diameter of the fuel particles X at this time is increased, the amount of soot and NOx generated can be reduced to almost zero. However, the amount of soot and NOx generated can be reduced to almost zero when the fuel vapor around the fuel particle X is burned at an appropriate time, and the combustion of the fuel vapor around the fuel particle X starts. Soot and NO when the time comes early
x will occur. Next, FIG.
This will be described with reference to FIG.

When the temperature of the intake gas in the combustion chamber 5 reaches approximately 270 ° C., the fuel vapor around the fuel particles X starts burning. When the combustion of the evaporated fuel around the fuel particles is started, a small peak of the heat generation rate appears as indicated by G in FIG. Next, a substantially constant clan angle Δθ of about 10 degrees after the start of combustion of the evaporated fuel around the fuel particles.
After that, all the fuel is ignited and burned, and thus a large peak of the heat generation rate appears as indicated by H in FIG.

As described above, the combustion indicated by G in FIG. 10A is started when the temperature of the intake gas in the combustion chamber 5 reaches approximately 270 ° C. Therefore, in the early stage of the compression stroke, the temperature of the intake gas in the combustion chamber 5 becomes approximately 270 ° C.
As shown in FIG. 10 (B), the combustion G of the fuel vapor around the fuel particles is started at an early stage of the compression stroke. However, when the start time of the combustion G is advanced in this way, the time of the combustion H of all the fuel is advanced, and as a result, soot and NOx are generated. Therefore, in the present invention, at the end of the compression stroke, in the operating state in which the temperature of the intake gas in the combustion chamber 5 becomes high, the temperature of the intake gas is reduced, and as a result, as shown in FIG. Combustion G is started after 10 degrees BTDC.

That is, in the embodiment shown in FIGS. 1 to 3, the EGR gas is supplied from the EGR passage 15 into the intake manifold 11 irrespective of the operation state of the engine. When the temperature of the intake gas in the combustion chamber 5 becomes high, the EGR gas is cooled by the cooling device 17, thereby lowering the temperature of the intake gas in the combustion chamber 5.

That is, as shown in the flowchart of FIG. 11, first, at step 60, it is determined whether or not the engine is operating under a high load, that is, the depression amount L of the accelerator pedal 27.
There whether large is determined than the set value L _0. L> L
When it is _0, the routine proceeds to step 61, where the electromagnetic switching valve 24 is turned on. At this time, the cooling water is supplied to the cooling water inlet 18 of the cooling device 17, and thus the EGR gas is
7 is cooled. On the other hand, when L ≦ L _0, the routine proceeds to step 62, where the electromagnetic switching valve 24 is turned off. At this time, the cooling water is returned through the bypass conduit 25 and the return conduit 20, and thus the cooling water
The cooling action of the GR gas is stopped.

As described above, in this embodiment, during the high load operation in which the intake gas temperature in the combustion chamber 5 becomes high at the end of the compression stroke, the EG
The R gas is cooled by the cooling device 17. as a result,
The temperature of the intake gas in the combustion chamber 5 at the end of the compression stroke is lowered, and thus, as shown in FIG. 10 (A), the combustion G can be started substantially at BTDC 10 degrees before the top dead center of the compression stroke.

FIG. 12 shows another embodiment. Referring to FIG. 12, in this embodiment, a cooling device 70 for cooling intake air between the intake manifold 11 and the air cleaner 12 is provided.
Is arranged. The cooling device 70 has a refrigerant inlet 71 and a refrigerant outlet 72, and the refrigerant inlet 71 and the refrigerant outlet 72 are connected to a cooler 73. In this embodiment, first, as shown in the flowchart of FIG.
That is, if the amount of depression L of the accelerator pedal 27 is larger than the set value L ₀ is determined. When L> L ₀ is cooler 73 is actuated proceeds to step 81. At this time, the refrigerant is supplied to the refrigerant inlet 71 of the cooling device 70,
Thus, the intake air is cooled by the cooling device 70.
On the other hand, when L ≦ L ₀ , the routine proceeds to step 82, where the cooler 73 is stopped, and the cooling operation of the intake air by the cooling device 70 is stopped.

As described above, in this embodiment, at the end of the compression stroke, the intake air is cooled by the cooling device 70 during the high load operation in which the intake gas temperature in the combustion chamber 5 becomes high. As a result, the temperature of the intake gas in the combustion chamber 5 at the end of the compression stroke is reduced, and thus, as shown in FIG. 10 (A), the combustion G is started approximately 10 degrees after the BTDC before the compression top dead center. Can be.

As can be understood from the above description, the fuel particles are uniformly dispersed by injecting the fuel at an early stage of the compression stroke, and the diameter of the fuel particles at this time is made larger than 500 μm. As shown in FIG. 10 (A), the start timing of combustion G
By starting after 10 degrees TDC, the generation of soot and NOx can be prevented. However, in practice, it is difficult to uniformly disperse the fuel particles in the combustion chamber 5. If the fuel particles cannot be uniformly dispersed in the combustion chamber 5, the start timing of the combustion G of the evaporated fuel around the fuel particles is considered. As shown in FIG. 10 (A), BTD almost before compression top dead center
Even if it is started after C10 degrees, soot and NOx are generated by the combustion G.

That is, since the fuel particles X are not uniformly dispersed in the combustion chamber 5 as described above, the dispersion density of the fuel particles X actually varies in the combustion chamber 5. That is,
In the combustion chamber 5, a region where the dispersion density of the fuel particles X is high as shown in FIG. 8B and a region where the dispersion density is low as shown in FIG. However, if a region where the dispersion density of the fuel particles X is high in the combustion chamber 5 as shown in FIG. 8B, soot NOx is generated in this region. In this case, in order to prevent the generation of soot and NOx, it is only necessary to suppress the oxidation reaction of the evaporated fuel around the fuel particles X. For that purpose, the density of oxygen around the fuel particles X should be reduced. It will be good.

Therefore, in the embodiment according to the present invention, the fuel particles X
In order to reduce the density of the surrounding oxygen, the exhaust gas recirculation rate (EGR gas amount / (EGR gas amount + intake air amount)), that is, the EGR rate is set to approximately 40% or more. That is, when the EGR rate is set to approximately 40% or more, the density of oxygen around the fuel particles X decreases, and thus the oxidation reaction of the fuel vapor around the fuel particles X is suppressed. In this way, the thermal decomposition of hydrocarbons in the fuel particles X is suppressed, and the amount of hydrogen molecules H ₂ and carbon C generated is reduced, so that the dispersion density of the fuel particles X is reduced as shown in FIG. Is high, the generation of soot and NOx is prevented.

FIG. 14 shows NOx during engine high load operation.
The relationship between the generation amount and the EGR rate is shown. From FIG.
It can be seen that when the EGR rate becomes about 40% or more, the NOx generation amount becomes extremely small, and when the EGR rate becomes about 50%, the NOx generation amount becomes almost zero. It has been found that soot becomes almost zero when the EGR rate becomes almost 50%.

Therefore, in the embodiment according to the present invention, as shown in FIG. 15, the EGR rate is maintained at almost 50% for all engine loads. FIG.
5A shows the relationship between the engine load, that is, the depression amount L of the accelerator pedal 27 and the EGR rate, and FIG. 15B shows the relationship between the excess air ratio λ and the depression amount L of the accelerator pedal 27. ing. As described above, in this embodiment, FIG.
As shown in (A), the EGR rate is always maintained at about 50% regardless of the depression amount L of the accelerator pedal 27.

In this embodiment, as can be seen from FIG. 15 (B), the excess air ratio λ is set to approximately 1.0 during engine high load operation, that is, when the depression amount L of the accelerator pedal 27 is large. In other words, in other words, the air-fuel ratio, which is the ratio between the amount of fuel and the amount of air in the combustion chamber 5, is substantially the stoichiometric air-fuel ratio. More specifically, at this time, the excess air ratio λ is controlled so that 2 to 3% of the air becomes excessive. When the excess air ratio λ is set to approximately 1.0, the density of oxygen around the fuel particles X is further reduced, and thus the oxidation reaction of the evaporated fuel around the fuel particles X is further suppressed. . As a result, the generation of soot and NOx can be further suppressed.

In this embodiment, the accelerator pedal 27
15B, the excess air ratio λ increases as the depression amount L of the accelerator pedal 27 decreases, as shown in FIG. That is, since the NOx generation amount decreases as the engine load decreases, the excess air ratio λ increases in this embodiment as the engine load decreases.

In this embodiment, the EGR rate is approximately 50%.
Although set to percent, the EGR rate can be set to any value between approximately 40 percent and approximately 60 percent as shown in FIG. FIG.
Shows another embodiment. In this embodiment, FIG.
As shown in (B), the excess air ratio λ is controlled to approximately 1.0 irrespective of the depression amount L of the accelerator pedal 27, that is, the air-fuel ratio is controlled to approximately the stoichiometric air-fuel ratio.
As shown in FIG. 6 (A), the EGR rate increases as the depression amount L of the accelerator pedal 27 decreases.

In these embodiments, the excess air ratio λ is
The fuel injection amount is determined so that the excess air ratio shown in FIG. 16B or FIG. 16B is obtained. The fuel injection amount Q is a function of the depression amount L of the accelerator pedal 27 and the engine speed N. 17 is stored in the ROM 32 in advance in the form of a map shown in FIG. Further, as shown in FIG. 18B, the fuel injection start timing θS is advanced as the depression amount L of the accelerator pedal 27 increases, that is, as the engine load increases. This fuel injection start timing θS is determined by the accelerator pedal 2
7 as a function of the stepping amount L and the engine speed N in FIG.
It is stored in the ROM 32 in advance in the form of a map shown in FIG.

On the other hand, in this embodiment, the exhaust manifold 13
The EGR gas flows from the EGR passage 15 into the intake manifold 11 due to the pressure difference between the back pressure in the intake manifold 11 and the pressure in the intake manifold 11.
The EGR gas amount at this time is controlled by the EGR control valve 16. The EG necessary for setting the EGR rate to the EGR rate shown in FIG. 15 (A) or FIG. 16 (A)
The opening degree of the R control valve 16, that is, the duty ratio of the drive pulse of the EGR control valve 16 is obtained in advance by an experiment,
The target duty ratio DUO is stored in the ROM 32 in advance in the form of a map shown in FIG. 19 as a function of the depression amount L of the accelerator pedal 27 and the engine speed N.

Further, based on the output signal of the air-fuel ratio sensor 26, the target duty ratio DU is set so that the excess air ratio λ becomes the target excess air ratio shown in FIG. 15 (B) or FIG. 16 (B).
O is corrected by the correction coefficient K. That is, the air-fuel ratio sensor 26 generates a current I according to the excess air ratio λ as shown in FIG. 20, and the actual air-fuel ratio is detected from the current value I. On the other hand, the target excess air ratio λO shown in FIGS. 15B and 16B is stored in the ROM 32 in advance in the form of a map shown in FIG. 21 as a function of the depression amount L of the accelerator pedal 27 and the engine speed N. Have been. In the embodiment shown in FIG. 16, the target excess air ratio λO is substantially 1.0 regardless of the operating state of the engine.

Next, the operation control routine shown in FIG. 22 will be described. Referring to FIG. 22, first, at step 100, the fuel injection amount Q is obtained from the map shown in FIG.
Is calculated. Next, in step 101, FIG.
Is calculated from the map shown in FIG.
Next, at step 102, the fuel injection completion timing θE is calculated based on the fuel injection amount Q, the fuel injection start timing θS, and the engine speed N.
Is calculated, and the fuel injection from the fuel injection valve 10 is performed based on the fuel injection start timing θS and the fuel injection completion timing θE.

Next, at step 103, the target duty ratio DUO is calculated from the map shown in FIG. 19, and then at step 104, the target excess air ratio λO is calculated from the map shown in FIG. Next, at step 105, the current excess air ratio λ is calculated from the output signal of the air-fuel ratio sensor 26. Next, at step 106, the current excess air ratio λ
Is larger than the target excess air ratio λO. When λ> λ0, the routine proceeds to step 107, where a fixed value α is added to the correction coefficient K, and then the routine proceeds to step 109. On the other hand, when λ ≦ λO, the routine proceeds to step 108, where the constant α is subtracted from the correction coefficient K, and then the routine proceeds to step 109. In step 109, the final duty ratio DU (= K · DUO) is calculated by multiplying the target duty ratio DUO by the correction coefficient K, and the EGR control valve 16 is controlled based on this duty ratio DU.

That is, when the actual excess air ratio λ becomes larger than the target excess air ratio λO, the duty ratio DU is increased. As a result, the opening degree of the EGR control valve 16 is increased, so that the EGR rate is increased, and thus the excess air ratio λ is reduced. On the other hand, when the actual excess air ratio λ becomes smaller than the target excess air ratio λO, the duty ratio DU is reduced. As a result, the opening degree of the EGR control valve 16 is reduced, so that the EGR rate is reduced, and thus the excess air ratio λ is increased. In this way, the excess air ratio λ is controlled to the target excess air ratio λO while maintaining the EGR rate substantially at the target EGR rate. FIG.
Shows still another embodiment. In this embodiment, a large amount of EGR
A throttle valve 121 is arranged in an intake duct 120 between the intake manifold 11 and the air cleaner 12 so that the gas can be easily recirculated. The throttle valve 121 is controlled by a step motor 122 to an opening degree corresponding to the operating state of the engine. However, in the case where the throttle valve 121 is provided in this manner, if the opening degree of the throttle valve 121 becomes small, the pressure in the combustion chamber 5 at the start of compression decreases, so the temperature of the intake gas in the combustion chamber 5 at the end of the compression stroke decreases. Resulting in. In particular, at the time of engine low load operation with a small amount of generated heat, the temperature of the intake gas in the combustion chamber 5 at the end of the compression stroke is significantly reduced, and good combustion cannot be obtained.

Therefore, in this embodiment, an electric heater device 123 for heating the intake air is installed in the intake duct 120, and the electric heater device 123 is operated when the engine load is low. That is, whether the time of engine high load operation, first, at step 130, as shown in the flowchart of FIG. 24, i.e. whether the amount of depression L of the accelerator pedal 27 (FIG. 1) is larger than the set value L ₁ not Is determined. L> step at the time of the L ₁ 13
Proceeding to 1, the electromagnetic switching valve 24 is turned on. At this time, the cooling water is supplied to the cooling water inlet 18 of the cooling device 17, and thus the EGR gas is cooled by the cooling device 17.
This electromagnetic switching valve 24 proceeds to step 132 when L ≦ L ₁ is turned off with respect. At this time, the cooling water is returned through the bypass conduit 25 and the return conduit 20,
Thus, the cooling operation of the EGR gas by the cooling device 17 is stopped.

Next, at step 133, it is determined whether or not the engine is in low load operation, that is, the depression amount L of the accelerator pedal 27 is determined.
Is smaller than the set value L ₂ (L ₂ <L ₁ ). When L <the L ₂ is an electric heater 123 is turned on proceeds to step 134. As a result, the intake air is heated, and thus the temperature of the intake gas in the combustion chamber 5 at the end of the compression stroke is increased. On the other hand, L ≧
When L ₂ is the electric heater 12 proceeds to step 135
3 is turned off.

FIG. 25 shows still another embodiment. In this embodiment, the temperature of the intake gas in the combustion chamber 5 at the end of the compression stroke during the low engine load operation is increased by controlling the opening timing of the intake valve 6 or the exhaust valve 8. In this embodiment, a phase control device 141 for the camshaft 140 for driving the intake valve 6 and a phase control device 143 for the camshaft 142 for driving the exhaust valve 8 are provided.
The valve opening periods of the intake valve 6 and the exhaust valve 8 are controlled by 1,143.

[0052] That is, whether the time of engine high load operation, first, at step 150, as shown in the flowchart of FIG. 26, i.e. than the amount of depression L is set value L ₁ of the accelerator pedal 27 (FIG. 1) It is determined whether it is larger. When L> L ₁ is the electromagnetic switching valve 24 is turned on proceeds to step 151. At this time, the cooling water is
The EGR gas is supplied to the cooling water inlet 18 of the cooling device 7 and thus cooled by the cooling device 17. On the other hand, L ≦
The electromagnetic switching valve 24 proceeds to step 152 when the L ₁
Is turned off. At this time, the cooling water is returned through the bypass conduit 25 and the return conduit 20, and the cooling operation of the cooling device 17 for cooling the EGR gas is stopped.

Next, at step 153, it is determined whether or not the engine is in low load operation, that is, the depression amount L of the accelerator pedal 27 is determined.
Is smaller than the set value L ₂ (L ₂ <L ₁ ). When L ≧ L _2, the routine proceeds to step 155, where the opening timing of the intake valve 6 and the exhaust valve 8 is controlled to the normal opening timing shown in FIG. Or hand L <opening timing of the intake valve 6 and the exhaust valve 8 proceeds to step 154 when the L ₂ is open period shown in FIG. 27 (B) 2
It is controlled during the valve opening period shown in FIG.

That is, in the case shown in FIG. 27B, the opening period of the exhaust valve 8 is advanced. If the valve opening period of the exhaust valve 8 is advanced, the amount of high temperature burned gas remaining in the combustion chamber 5 increases, so that the temperature of the intake gas in the combustion chamber 5 increases. On the other hand, in the case shown in FIG. 27C, the opening period of the intake valve 6 is delayed. If the opening period of the intake valve 6 is delayed, the intake air suddenly flows into the combustion chamber 5 when the intake valve 6 opens, and at this time, the inertial supercharging action of the intake air causes the internal combustion chamber 5 Since the intake air is adiabatically compressed, the temperature of the intake gas in the combustion chamber 5 is increased.

In addition, the compression ratio in the combustion chamber 5 may be controlled to increase the compression ratio in the combustion chamber 5 during low-load operation of the engine.

[0056]

According to the present invention, the amounts of soot and NOx generated can be reduced to almost zero.

[Brief description of the drawings]

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

FIG. 2 is a side sectional view of a compression ignition type internal combustion engine.

FIG. 3 is a bottom view of the cylinder head of FIG. 2;

FIG. 4 is a side sectional view of the fuel injection valve.

FIG. 5 is an enlarged side sectional view of a front end portion of the fuel injection valve.

FIG. 6 is a diagram showing a pressure change in a combustion chamber caused only by a compression action of a piston.

FIG. 7 is a diagram showing a boiling point and a change in temperature of fuel particles.

FIG. 8 is a diagram showing a distribution of fuel particles.

FIG. 9 is a diagram showing a distribution of fuel particles.

FIG. 10 is a diagram showing a heat generation rate.

FIG. 11 is a flowchart for performing cooling control.

FIG. 12 is an overall view of a compression ignition type internal combustion engine showing another embodiment.

FIG. 13 is a flowchart for performing cooling control.

FIG. 14 is a diagram showing the relationship between the NOx generation amount and the EGR rate.

FIG. 15 is a diagram showing an EGR rate and an excess air rate.

FIG. 16 is a diagram showing an EGR rate and an excess air rate.

FIG. 17 is a diagram showing a map of a fuel injection amount.

FIG. 18 is a diagram showing a fuel injection start timing.

FIG. 19 is a diagram showing a map of a target duty ratio.

FIG. 20 is a diagram showing a current generated by an air-fuel ratio sensor.

FIG. 21 is a diagram showing a target excess air ratio.

FIG. 22 is a flowchart for controlling engine operation.

FIG. 23 is an overall view of a compression ignition type internal combustion engine showing still another embodiment.

FIG. 24 is a flowchart for performing intake gas temperature control.

FIG. 25 is an overall view of a compression ignition type internal combustion engine showing still another embodiment.

FIG. 26 is a flowchart for performing intake gas temperature control.

FIG. 27 is a diagram showing lift curves of an intake valve and an exhaust valve.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 11 ... Intake manifold 13 ... Exhaust manifold 15 ... EGR passage 16 ... EGR control valve 17 ... Cooling device

──────────────────────────────────────────────────続 き Continued on the front page (51) Int.Cl. ⁷ Identification code FI F02D 13/02 F02D 13/02 J 21/08 301 21/08 301D F02M 25/07 550 F02M 25/07 550G 570 570F 580 580E 58) Field surveyed (Int.Cl. ⁷ , DB name) F02D 41/02 301 F02D 41/02 380 F02D 21/08 301 F02M 25/07 550

Claims (2)

(57) [Claims] 1. Injecting fuel into a combustion chamber during a compression stroke or an intake stroke approximately 60 degrees before compression top dead center, and determining the average particle diameter of the injected fuel at this time so that the temperature of the fuel particles is substantially above the compression stroke. After the dead center or the compression top dead center, the particle size must reach the boiling point of the main fuel component determined by the pressure at that time,
In-cylinder injection that prevents fuel evaporation due to boiling from the fuel particles until it almost reaches the compression top dead center after injection, and causes the fuel in the fuel particles to boil and evaporate after the compression top dead center to ignite and burn the fuel. Compression top dead
Temperature lowering means for lowering the temperature of the intake gas sucked into the combustion chamber at the end of the compression stroke in order to start combustion after 10 degrees before the engine load, and when the engine load is higher than a predetermined load. An in-cylinder injection type in which the temperature lowering means lowers the temperature of the intake gas at the end of the compression stroke and stops the temperature lowering action of the intake gas by the temperature lowering means when the engine load is lower than a predetermined load. Internal combustion engine. 2. A temperature increasing means for increasing the temperature of the intake gas sucked into the combustion chamber at the end of the compression stroke. When the engine load is lower than a predetermined set value, the temperature is increased by the temperature increasing means. 2. The in-cylinder injection according to claim 1, wherein the temperature at the end of the gas compression stroke is raised, and when the engine load is higher than a predetermined set value, the action of raising the temperature of the intake gas by the temperature raising means is stopped. Type internal combustion engine.