JP3978965B2 - Combustion control device for internal combustion engine - Google Patents

Description

[0001]
BACKGROUND OF THE INVENTION
The present invention relates to an internal combustion engine that performs premixed compression self-ignition combustion, and more particularly to an internal combustion engine that performs compression self-ignition combustion in a wide operating range by stabilizing the combustion timing by optimizing the mixture distribution in the cylinder. The present invention relates to a combustion control device.
[0002]
[Prior art]
In order to improve the thermal efficiency of a gasoline engine, a method is known in which the air-fuel mixture is leaned to reduce pump loss and increase the specific heat ratio of the working gas to increase the theoretical thermal efficiency. However, in the conventional spark ignition engine, when the air-fuel ratio is made lean, the combustion period becomes longer and the combustion stability deteriorates. For this reason, there is a limit to the lean air-fuel ratio.
[0003]
As a technique for leaning the air-fuel ratio while avoiding such deterioration in combustion stability, a two-stroke cycle engine that causes premixed compression self-ignition combustion is disclosed in JP-A-7-71279. In premixed compression self-ignition combustion, combustion reactions occur from multiple positions in the combustion chamber, so even when the air-fuel ratio is lean, the combustion period is not prolonged compared to spark ignition and stable even at a leaner air-fuel ratio. Combustion becomes possible. Further, since the air-fuel ratio is lean, the combustion temperature is lowered, and NOx (nitrogen oxide) can be greatly reduced.
[0004]
As a second conventional technique, a compression ignition type internal combustion engine is disclosed in which fuel is divided into two portions and supplied into a cylinder in one cycle as disclosed in JP-A-11-236848. In this conventional example, in order to moderate the combustion of the diesel engine and improve the knocking limit and reduce the generation amount of NOx and soot, the first injection of 30% or less of the total injection amount at a predetermined injection timing before the compression top dead center. Next, the remaining fuel is injected at the second injection timing almost near the compression top dead center.
[0005]
[Problems to be solved by the invention]
However, since the first conventional example has a normal two-stroke cycle engine configuration, there is no intake valve and exhaust valve for controlling gas exchange, and unburned gas blows out, resulting in deterioration of fuel consumption. In addition, since the expansion stroke is an expansion / exhaust stroke in which exhaust is performed in the latter half because of the need to perform gas exchange, work due to expansion of the combustion gas cannot be taken out sufficiently, which makes it difficult to perform high-load operation. there were.
[0006]
When the second fuel injection amount is increased from the first time as in the second prior art, the thin air-fuel mixture generated by the first injection is a fuel with very good ignitability such as light oil. However, in the case of a fuel with poor ignitability such as gasoline, a large amount of unburned HC is discharged.
[0007]
Further, since the second injection amount is relatively large, there is a problem that the combustion temperature becomes high and it is difficult to reduce NOx.
[0008]
The present invention has been made in view of such problems, and its object is to control combustion of an internal combustion engine capable of improving the fuel consumption by expanding the operating range by compression self-ignition combustion while avoiding knocking and combustion instability. Is to provide a device.
[0009]
Another object of the present invention is to provide a clean internal combustion engine that reduces unburned HC (hydrocarbon) and NOx by compression self-ignition combustion.
[0010]
[Means for Solving the Problems]
In order to solve the above-mentioned problem, the invention according to claim 1 performs the first and second fuel injections for each combustion cycle from the fuel injection valve for directly injecting fuel into the cylinder, and the piston is compressed by the upward movement of the piston. In a combustion control apparatus for an internal combustion engine that ignites and burns a high-temperature and high-pressure air-fuel mixture, a first air-fuel mixture region that does not ignite near the compression top dead center is formed by the first fuel injection, and compression is performed by the second fuel injection A second gas mixture region that ignites near the top dead center is formed in the vicinity of the central axis of the combustion chamber, the first gas mixture region is formed in the outer periphery of the second gas mixture region, and the vicinity of the combustion chamber wall is the fuel. The gist is that the amount of injection in the second fuel injection is less than the first fuel injection amount at least near the knock limit.
[0013]
In order to solve the above-mentioned problem, according to a second aspect of the present invention, in the combustion control device for an internal combustion engine according to the first aspect , the injection amount in the second fuel injection is about the total injection amount at least near the knock limit. The gist is that it is 40% or less.
[0018]
In order to solve the above-mentioned problem, an invention according to claim 3 is the combustion control device for an internal combustion engine according to claim 1 or 2, wherein the excess air ratio between the first mixture region and the second mixture region is set. The difference is that it is about 1.0 or less at least near the knock limit.
[0019]
In order to solve the above-mentioned problems, a fourth aspect of the present invention is the combustion control device for an internal combustion engine according to any one of the first to third aspects, wherein the ignition timing of the first mixture region is after the end of the compression stroke. Thus, the gist is to control the second fuel injection timing.
[0020]
In order to solve the above-mentioned problem, a fifth aspect of the present invention provides the combustion control apparatus for an internal combustion engine according to any one of the first to fourth aspects, wherein the difference in the excess air ratio according to a decrease in engine load. Or reducing the injection amount in the second fuel injection, or delaying the first fuel injection timing.
[0021]
【The invention's effect】
According to the first aspect of the present invention, the first and second fuel injections are performed for each combustion cycle from the fuel injection valve that injects the fuel directly into the cylinder, and the high-temperature and high-pressure compressed by the upward movement of the piston is performed. In a combustion control apparatus for an internal combustion engine that ignites and burns an air-fuel mixture, a first air-fuel mixture region that does not ignite near the compression top dead center is formed by the first fuel injection, and a compression top dead center is formed by the second fuel injection. A second gas mixture region that ignites in the vicinity is formed in the vicinity of the center axis of the combustion chamber, the first gas mixture region is formed in the outer periphery of the second gas mixture region, and there is almost no fuel in the vicinity of the combustion chamber wall. Since the air amount is set so that the injection amount in the second fuel injection is smaller than the first fuel injection amount at least in the vicinity of the knock limit, the second mixture region is near the compression top dead center. The second gas mixture region ignited reliably and ignited To spread gradually combustion to the first mixture region Luo non-ignition, thereby improving the self-ignition of the internal combustion engine, there is an effect that knocking can be improved preventing acoustic vibrations characteristic.
In addition, since the second air-fuel mixture region that ignites near the compression top dead center is formed near the center axis of the combustion chamber, it is ignited from the vicinity of the center portion of the combustion chamber and isotropically spread to the periphery of the combustion chamber. As a result, the combustion pressure applied to the gas can be made uniform and vibrations can be further reduced.
[0022]
Further, the first mixture region is formed by the first fuel injection, and the second mixture region is formed by the second fuel injection in addition to the first fuel injection. Therefore, there is an effect that both the mixture regions can be formed only by controlling the fuel injection timing and the number of injections without using a complicated device for forming the first and second mixture regions.
[0023]
In addition, since the injection amount in the second fuel injection is smaller than the first fuel injection amount at least near the knock limit, the space of the portion where the fuel concentration is high and the fuel is burned at high temperature is reduced, and the NOx emission amount is reduced. In addition to reducing the amount of particulate matter, the amount of particulate matter discharged can be reduced.
Furthermore, since there is no fuel in the low-temperature portion near the cylinder wall, the fuel is burned out before the flame reaches the vicinity of the cylinder wall, and the HC emission amount can be reduced to almost zero.
[0024]
According to the invention described in claim 2 , in addition to the effect of the invention described in claim 1 , the injection amount in the second fuel injection is about 40% or less of the total injection amount at least near the knock limit. As a result, the amount of NOx and particulate matter discharged can be further reduced.
[0029]
According to the invention described in claim 3 , in addition to the effect of the invention described in claim 1 or 2, the difference in the excess air ratio between the first mixture region and the second mixture region is: By setting it to about 1.0 or less at least near the knock limit, NOx can be further greatly reduced.
[0030]
According to the invention described in claim 4 , in addition to the effects of the invention described in claims 1 to 3 , the second time is set so that the ignition timing of the first mixture region comes after the end of the compression stroke. Since the fuel injection timing is controlled, combustion is performed in the expansion stroke in which the volume of the combustion chamber is expanded, so that the in-cylinder pressure increase rate that causes knocking can be effectively reduced.
[0031]
According to the fifth aspect of the present invention, in addition to the effects of the first to fourth aspects of the invention, the difference in the excess air ratio is reduced or the first time according to the decrease in the engine load. Since the fuel injection timing is retarded or the injection amount in the second fuel injection is reduced, it is possible to reduce both HC and NOx at high load and HC and NOx at low load. it can.
[0032]
DETAILED DESCRIPTION OF THE INVENTION
[First Embodiment]
Next, embodiments of the present invention will be described in detail with reference to the drawings.
FIG. 1 is a system configuration diagram showing a configuration of a first embodiment in which a combustion control device for an internal combustion engine according to the present invention is applied to a gasoline engine.
[0033]
In the present invention, the first air-fuel mixture region 6 that does not ignite near the compression top dead center is formed by the first fuel injection, and the second air-fuel mixture ignited near the compression top dead center by the second fuel injection. It is characteristic that the region 7 is formed.
[0034]
The engine body 10 in the figure includes a cylinder 11, a cylinder head 12, a piston 13, an intake port 14, an intake valve 15 that opens and closes the intake port 14, an exhaust port 16, and an exhaust valve that opens and closes the exhaust port 16. 17 and a fuel injection valve 18 for injecting fuel directly into the cylinder.
[0035]
In the figure, reference numeral 6 denotes a first mixture region, reference numeral 7 denotes a second mixture region, and reference numeral 8 denotes an air layer near the combustion chamber wall.
[0036]
An electronic control unit (hereinafter abbreviated as ECU) 1 that controls the engine body 10 determines operating conditions based on a crank angle sensor signal and an accelerator opening sensor signal, and sets the number of fuel injections once or twice. An operation condition determination unit 2, a first fuel injection control unit 3 that forms the first mixture region 6 by controlling the first fuel injection according to the determination of the operation condition determination unit 2, and the operation condition determination unit 2 According to the signal from the second fuel injection control unit 4 that forms the second mixture region 7 by controlling the second fuel injection according to the determination, and the first fuel injection control unit 3 and the second fuel injection control unit 4 And a fuel injection valve drive unit 5 that outputs a drive signal for the fuel injection valve 18.
[0037]
In addition, although the component of ECU1 can also be comprised with a hard-wired logic circuit, it can also be implement | achieved as a microcomputer and its control program.
[0038]
FIG. 2 is an example of a plan view of a main part of the engine body 10. In FIG. 2, the engine body 10 includes two intake ports 14, intake valves 15, exhaust ports 16, and exhaust valves 17, improving the intake / exhaust efficiency at high rotation and high load, and the mass of the valve operating system. Is reduced. One of the intake ports 14 is provided with a swirl control valve 19, and by closing the swirl control valve 19, a lateral swirl flow (swirl) is formed in the cylinder so that mixing of fuel and intake air can be promoted. Yes.
[0039]
However, the present invention does not necessarily require that the intake and exhaust valves are multi-valves, and the swirl control valve is not essential.
[0040]
FIG. 3 is a longitudinal sectional view showing the shape of the distal end portion of a holocorn injection valve 20 which is an example of the fuel injection valve 18 used in the present embodiment. The hollow cone injection valve 20 is provided with one injection hole 22 in the center of the valve body tip 21 in the injection valve axial direction and a needle valve 23 that opens and closes the injection hole 22. When a fuel injection pulse is energized to a drive coil (not shown), the needle valve 23 is lifted upward in the figure, and the nozzle 22 is opened to inject fuel. The valve lift amount of the needle valve 23 may be variable according to a desired injection period.
[0041]
The fuel injected from the injection port 22 spreads in a conical surface shape coaxial with the axis of the holocon injection valve 20. The hollow cone injector 20 can be used at a relatively low fuel pressure.
[0042]
If the fuel injection timing using the hollow cone injection valve 20 is relatively early during the intake stroke or compression stroke, the injected fuel can fly easily because the in-cylinder pressure is not so high. It is distributed over a relatively wide area. On the contrary, if the fuel injection timing is relatively late during the compression stroke, the in-cylinder pressure is high, so that the injected fuel becomes difficult to fly and does not spread too much.
[0043]
FIG. 4 is a longitudinal sectional view (a) and a plan view (b) showing the shape of the distal end portion of a multi-injection injection valve 30 which is another example of the fuel injection valve 18 used in the present embodiment. The multi-injection injection valve 30 has one central injection port 32 in the injection valve axial direction at the center of the valve body tip 31 and four sides oriented around the central injection injection port 32 in a direction inclined with respect to the injection valve axial direction. A direction nozzle 33 and a needle valve 34 are provided. The valve lift amount of the needle valve 34 may be variable according to a desired injection period. The side injection holes 33 are preferably arranged rotationally symmetrically around the fuel injection valve shaft or the cylinder axis, and the number thereof is not necessarily limited to four but may be a plurality, preferably three or more.
[0044]
When a fuel injection pulse is applied to a drive coil (not shown), the needle valve 34 is lifted upward in the figure, the central injection port 32 and the plurality of side injection ports 33 are opened, and fuel is injected. The fuel injected from the central injection port 32 and the side injection port 33 spreads in a conical shape coaxial with the axis of the multi-injection injection valve 30. The multi-injection injection valve 30 is used at a relatively high fuel pressure, the particle size of the injected fuel is reduced, and particulate matter in the exhaust can be greatly reduced.
[0045]
If the fuel injection timing using the multi-injection valve 30 is relatively early during the intake stroke or compression stroke, the injected fuel can fly easily because the cylinder pressure is not so high. It is distributed over a relatively wide area. On the contrary, if the fuel injection timing is relatively late during the compression stroke, the in-cylinder pressure is high, so that the injected fuel becomes difficult to fly and does not spread too much.
[0046]
FIG. 5 is a graph showing an example of the in-cylinder fuel distribution at the time of high load formed by the fuel injection of the first embodiment. In the figure, the horizontal axis indicates the distance from the cylinder central axis (combustion chamber central axis), and the vertical axis indicates the equivalence ratio. The fuel distribution around the cylinder axis is a uniform distribution, that is, a distribution obtained by rotating FIG. 5 about the vertical axis.
[0047]
As shown in the figure, when the load is high, a second air-fuel mixture region 7 is formed near the center axis of the cylinder that has a high equivalence ratio and ignites near the compression top dead center. A first air-fuel mixture region 6 that does not ignite near the point is formed, and an air layer 8 with almost no fuel is formed near the cylinder wall.
[0048]
The first mixture region 6 is formed by being injected in the intake stroke or the compression stroke by the control of the first fuel injection control unit 3, and the second mixture region 7 is controlled by the control of the second fuel injection control unit 4. It is formed by fuel injection after the first fuel injection in the compression stroke until the compression top dead center. In order to ensure the formation of the air layer 8, it is desirable to perform the first fuel injection after the latter half of the intake stroke.
[0049]
FIG. 6 is a graph showing an example of the in-cylinder fuel distribution at the time of low load formed by the fuel injection of the first embodiment. In the figure, the horizontal axis indicates the distance from the cylinder central axis (combustion chamber central axis), and the vertical axis indicates the equivalence ratio. The fuel distribution around the cylinder axis is a uniform distribution, that is, a distribution obtained by rotating FIG. 6 about the vertical axis.
[0050]
As shown in the figure, at the time of low load, for example, one air-fuel mixture region 9 is formed from the vicinity of the cylinder central axis to about ½ of the bore radius, and an air layer 8 is provided outside the cylinder wall.
[0051]
The equivalence ratio of the air-fuel mixture region 9 at the time of low load is higher than that of the first air-fuel mixture region 6 at the time of high load and slightly lower than that of the second air-fuel mixture region 7. Formed by control. As a result, the ignitability at the time of low load is ensured, the fuel is burned out before the flame reaches the low-temperature cylinder wall, and the discharge of unburned HC can be made almost zero.
[0052]
FIG. 7 is a graph showing the NOx emission amount change when the in-cylinder volume ratio of the second mixture region is changed at high load. When the volume ratio of the second mixture region near the center axis of the combustion chamber is increased gradually from 0% to 100% of the combustion chamber volume at the compression top dead center at high load, the NOx emission increases as the volume ratio increases. . This is because the volume ratio of the region with a high equivalence ratio, in other words, the region with a high fuel concentration increases, and the volume ratio of the region with a high combustion temperature increases.
[0053]
FIG. 8 is a graph showing changes in the HC emission amount when the in-cylinder volume ratio of the second mixture region is changed at high load.
[0054]
When the volume ratio of the second gas mixture region near the center axis of the combustion chamber is increased gradually from 0% to 100% of the combustion chamber volume at the compression top dead center at the time of high load, the second gas mixture having good ignitability is initially obtained. Since there is almost no region, the ignitability is poor and the HC emission amount is large, but the ignitability improves as the volume ratio increases, so the HC emission amount suddenly decreases to almost zero. And since the volume ratio exceeds about 20%, HC discharge | emission amount increases as the volume ratio increases. Moreover, the rate of increase becomes more rapid as the volume ratio approaches 100%.
[0055]
This is because when the volume ratio of the second gas mixture region having good ignitability increases, the gas mixture region having a high equivalence ratio and high fuel concentration extends to the vicinity of the combustion chamber wall. For this reason, it is because a fuel spread | diffuses to the piston clevis part and the combustion chamber wall surface in the immediate vicinity, and the fuel layer which contact | connects these low temperature parts becomes a quench layer (flame extinction layer).
[0056]
From the above FIGS. 7 and 8, the second fuel-air mixture region of at high load, it would be preferable that the volume of about 10 to 30% of the combustion chamber volume product at compression top dead center.
[0057]
FIG. 9 is a graph showing changes in NOx emission when the in-cylinder volume ratio of the second gas mixture region is changed at low load. When the volume ratio of the second mixture region near the center axis of the combustion chamber at low load is gradually increased from 0% to 100% of the combustion chamber volume at the compression top dead center , the NOx emission decreases as the volume ratio increases. In about 50%, NOx is almost lost.
[0058]
FIG. 10 is a graph showing changes in the HC emission amount when the in-cylinder volume ratio of the second mixture region is changed at low load. When the volume ratio of the second gas mixture region near the center axis of the combustion chamber at low load is gradually increased from 0% to 100% of the combustion chamber volume at the compression top dead center , the HC emission is about 0% to 50%. Although there is almost no amount, the amount of HC emission increases gradually from the point where it exceeds about 50%, with the increase in the volume ratio, and then suddenly increases later. This is because, according to the volume ratio of the second gas mixture region having good ignitability, the fuel diffuses to the immediate vicinity of the piston clevis portion and the combustion chamber wall surface, and the fuel layer in contact with these low temperature portions becomes a quench layer.
[0059]
9 and 10 described above , the NOx and HC emissions are reduced by setting the second gas mixture region to a volume of about 50% of the combustion chamber volume at the compression top dead center when the load is low. Can do.
[0060]
FIG. 11 is a graph illustrating the control of the fuel injection amount load by the first fuel injection control unit 3 and the second fuel injection control unit 4.
[0061]
In the first injection amount, the injection amount linearly decreases as the load decreases from the injection amount at the time of high load. The second injection amount linearly decreases as the load decreases from the injection amount at the time of high load, and is zero near the low load. The first injection amount at high load is more than half of the total injection amount of the first and second injections, preferably 60 to 80%, and the second injection amount is the remaining amount 20 to 40 %.
[0062]
As the load decreases from a high load to a low load, the second injection amount is decreased. Therefore, the excess air ratio in the first mixture region formed by the first fuel injection and the second fuel injection are formed. The difference from the excess air ratio in the second air-fuel mixture region is 1.0 or less at high load and decreases as the load decreases.
[0063]
FIG. 12 is a graph illustrating control of the fuel injection timing by the first fuel injection control unit 3 and the second fuel injection control unit 4 with respect to the load.
[0064]
The second injection timing is set to a certain time before the top dead center in the latter half of the compression stroke, and is preferably set so that the timing at which the second gas mixture region undergoes compression self-ignition is after the end of the compression stroke. The first injection timing is an advanced angle close to BDC at high load, but is retarded just before the second injection timing, preferably 60 degrees before the end of the compression stroke, as the load decreases.
[0065]
[Second Embodiment]
Next, a second embodiment of the combustion control apparatus for an internal combustion engine according to the present invention will be described with reference to FIGS.
[0066]
FIG. 13 is a system configuration diagram showing the configuration of a second embodiment in which the combustion control device for an internal combustion engine according to the present invention is applied to a gasoline engine. Since the configuration of the present embodiment is substantially the same as the configuration of the first embodiment shown in FIG. 1, the same reference numerals are given to the same components, and redundant description is omitted.
[0067]
In addition, the principal part top view of the engine main body 10 in this embodiment is the same as that of 1st Embodiment shown in FIG. 2, The fuel injection valve 18 used for this embodiment is shown in FIG.3 and FIG.4. This is the same as in the first embodiment.
[0068]
In the second embodiment, the shape of the first air-fuel mixture region 6 formed by the first fuel injection from the fuel injection valve 18 that directly injects fuel into the cylinder, and the second fuel injection formed by the second fuel injection. 2 is characterized in that the shape of the air-fuel mixture region 7 is different from that of the first embodiment.
[0069]
FIG. 14 is a graph showing an example of in-cylinder fuel distribution at the time of high load formed by fuel injection according to the second embodiment. In the figure, the horizontal axis indicates the distance from the cylinder central axis (combustion chamber central axis), and the vertical axis indicates the equivalence ratio. The fuel distribution around the cylinder axis is a uniform distribution, that is, a distribution obtained by rotating FIG. 14 about the vertical axis.
[0070]
As shown in the figure, when the load is high, the equivalence ratio is relatively low in the vicinity of the center axis of the cylinder, and the first mixture region 6 that does not ignite near the compression top dead center is formed. A second air-fuel mixture region 7 that is highly ignited near the compression top dead center is formed in a donut shape, and further, an air layer 8 in the vicinity of the cylinder wall is almost free of fuel.
[0071]
The first mixture region 6 is formed by being injected in the intake stroke or the compression stroke by the control of the first fuel injection control unit 3, and the second mixture region 7 is controlled by the control of the second fuel injection control unit 4. It is formed by fuel injection after the first fuel injection in the compression stroke until the compression top dead center. In order to ensure the formation of the air layer 8, it is desirable to perform the first fuel injection after the latter half of the intake stroke.
[0072]
FIG. 15 is a graph showing an example of the in-cylinder fuel distribution at the time of low load formed by the fuel injection of the second embodiment. In the figure, the horizontal axis indicates the distance from the cylinder central axis (combustion chamber central axis), and the vertical axis indicates the equivalence ratio. The fuel distribution around the cylinder axis is a uniform distribution, that is, a distribution obtained by rotating FIG. 15 about the vertical axis.
[0073]
As shown in the figure, at the time of low load, for example, the first air-fuel mixture region 6 is formed from the vicinity of the cylinder center axis to about 1/3 of the bore radius, and the second equivalence ratio is relatively high on the outer periphery. The air-fuel mixture region 7 is formed in a donut shape up to about 2/3 of the bore radius, and an air layer 8 is provided on the outside to the cylinder wall.
[0074]
When the load is low, the difference between the equivalence ratio of the first mixture region 6 and the equivalence ratio of the second mixture region 7 is set to be larger than that during the high load. As a result, the ignitability at the time of low load is ensured, the fuel is burned out before the flame reaches the low-temperature cylinder wall, and the discharge of unburned HC can be made almost zero.
[0075]
FIG. 16 is a graph showing the NOx emission amount change when the in-cylinder volume ratio of the first mixture region is changed, and shows almost the same tendency from the time of high load to the time of low load.
[0076]
When the volume ratio of the first gas mixture region near the combustion chamber central axis is gradually increased from 0% to 100% of the combustion chamber volume at the compression top dead center , the NOx emission is almost zero until the volume ratio is about 35%. However, the NOx emission increases as the volume ratio increases, reaches a maximum around 70%, and then decreases.
[0077]
This is because the volume ratio of the second gas mixture region that ignites near the compression top dead center decreases as the volume ratio of the first gas mixture region that does not ignite near the compression top dead center increases. This is because the fuel concentration in the mixture region increases, the combustion temperature rises, and the NOx emission amount increases.
[0078]
The NOx emission amount decreases from the point where the volume ratio of the first mixture region exceeds about 70% of the combustion chamber volume at the compression top dead center to 100%. This is because the volume ratio of the second gas mixture region decreases with an increase in the amount of gas and the second gas mixture region diffuses into the outer air layer by the time of combustion, thereby reducing the region that burns at high temperature. is there.
[0079]
FIG. 17 is a graph showing a change in the HC discharge amount when the in-cylinder volume ratio of the first mixture region is changed, and shows almost the same tendency from a high load to a low load.
[0080]
If the volume ratio of the first gas mixture region near the combustion chamber central axis is gradually increased from 0% to 100% of the combustion chamber volume at the compression top dead center, the first mixture that does not ignite near the compression top dead center at first. Since there is almost no gas region and the volume ratio of the second gas mixture region that ignites near the compression top dead center occupies most, the equivalent ratio of the second gas mixture region as a whole decreases, and its ignitability is poor and HC emissions Although the amount is large, as the volume ratio of the first gas mixture region increases, the ignitability of the second gas mixture region becomes better, so the HC emission amount suddenly decreases to almost zero.
[0081]
The HC emission amount becomes almost zero when the volume ratio of the first mixture region is about 20% to about 45% of the combustion chamber volume at the compression top dead center . Since the volume ratio of the first gas mixture region exceeds about 45%, the HC emission amount increases as the volume ratio increases. Moreover, the rate of increase becomes more rapid as the volume ratio approaches 100%.
[0082]
This is because, as the volume ratio of the first air-fuel mixture region spreading from the center of the combustion chamber to the surroundings increases, the second air-fuel mixture region having good ignitability existing in the outer peripheral portion, that is, the fuel having a high equivalence ratio A high-concentration mixture region extends to the vicinity of the combustion chamber wall. For this reason, it is because a fuel spread | diffuses to the piston clevis part and the combustion chamber wall surface in the immediate vicinity, and the fuel layer which contact | connects these low temperature parts becomes a quench layer (flame extinction layer).
[0083]
From the above FIGS. 16 and 17, in this second embodiment, the first mixture region of, preferably a volume of about 20% to about 40% of the combustion chamber volume product at compression top dead center It can be said.
[0084]
FIG. 18 is a graph illustrating the control of the fuel injection amount load by the first fuel injection control unit 3 and the second fuel injection control unit 4.
[0085]
In the first injection amount, the injection amount linearly decreases as the load decreases from the injection amount at the time of high load. The second injection amount increases linearly as the load decreases from the injection amount at the time of high load, and the increase rate is smaller than the decrease rate of the first injection amount.
[0086]
The first injection amount at high load is more than half of the total injection amount of the first and second injections, preferably 60 to 80%, and the second injection amount is the remaining amount 20 to 40 %.
[0087]
As the load decreases from a high load to a low load, the second injection amount is decreased. Therefore, the excess air ratio in the first mixture region formed by the first fuel injection and the second fuel injection are formed. The difference from the excess air ratio in the second air-fuel mixture region is 0 to 1.0 at a high load, and decreases as the load decreases.
[0088]
FIG. 19 is a graph illustrating the control of the fuel injection timing by the first fuel injection control unit 3 and the second fuel injection control unit 4 with respect to the load.
[0089]
The second injection timing is set to a certain time before the top dead center in the latter half of the compression stroke, and is preferably set so that the timing at which the second gas mixture region undergoes compression self-ignition is after the end of the compression stroke. The first injection timing is the intake stroke (not shown) or the first half of the compression stroke as shown in FIG.
[Brief description of the drawings]
FIG. 1 is a system configuration diagram showing a configuration of a first embodiment in which a combustion control device for an internal combustion engine according to the present invention is applied to a gasoline engine.
FIG. 2 is an example plan view of a main part of an engine body according to the embodiment.
FIG. 3 is a longitudinal sectional view showing a shape of a tip portion of a holocorn injection valve used in the embodiment.
FIG. 4 is a longitudinal sectional view (a) and a plan view (b) showing the shape of the tip of a multi-injection valve used in the embodiment.
FIG. 5 is a graph showing an example of in-cylinder fuel distribution at the time of high load formed by fuel injection according to the first embodiment.
FIG. 6 is a graph showing an example of in-cylinder fuel distribution at low load formed by fuel injection according to the first embodiment.
FIG. 7 is a graph showing the NOx emission amount when the in-cylinder volume ratio of the second mixture region is changed at high load in the first embodiment.
FIG. 8 is a graph showing the amount of HC discharged when the in-cylinder volume ratio of the second mixture region is changed at high load in the first embodiment.
FIG. 9 is a graph showing the NOx emission amount when the in-cylinder volume ratio of the second gas mixture region is changed at low load in the first embodiment.
FIG. 10 is a graph showing the amount of HC discharged when the in-cylinder volume ratio of the second mixture region is changed at low load in the first embodiment.
FIG. 11 is a graph illustrating control of the first and second fuel injection amounts with respect to load in the first embodiment.
FIG. 12 is a graph for explaining the control of the load at the first and second fuel injection timings in the first embodiment.
FIG. 13 is a system configuration diagram showing a configuration of a second embodiment in which a combustion control device for an internal combustion engine according to the present invention is applied to a gasoline engine.
FIG. 14 is a graph showing an example of in-cylinder fuel distribution at the time of high load formed by fuel injection according to the second embodiment.
FIG. 15 is a graph showing an example of in-cylinder fuel distribution at low load formed by fuel injection according to the second embodiment.
FIG. 16 is a graph showing the NOx emission amount when the in-cylinder volume ratio of the first mixture region is changed in the second embodiment.
FIG. 17 is a graph showing the HC discharge amount when the in-cylinder volume ratio of the first mixture region is changed in the second embodiment.
FIG. 18 is a graph illustrating control of the first and second fuel injection amounts with respect to load in the second embodiment.
FIG. 19 is a graph for explaining the control over the load at the first and second fuel injection timings in the second embodiment.
[Explanation of symbols]
1 Electronic control unit (ECU)
2 Operating condition determination unit 3 1st fuel injection control unit 4 2nd fuel injection control unit 5 Fuel injection valve drive unit 6 First mixture region 7 Second mixture region 8 Air layer 10 Engine body 11 Cylinder block 12 Cylinder Head 13 Piston 14 Intake port 15 Intake valve 16 Exhaust port 17 Exhaust valve 18 Fuel injection valve 19 Swirl control valve

Claims (5)

Combustion control of an internal combustion engine that performs first and second fuel injections for each combustion cycle from a fuel injection valve that directly injects fuel into a cylinder and ignites and burns a high-temperature and high-pressure air-fuel mixture compressed by the upward movement of the piston In the device
A first mixture region that does not ignite near the compression top dead center is formed by the first fuel injection, and a second mixture region that ignites near the compression top dead center by the second fuel injection is the center of the combustion chamber. Formed near the shaft, the first mixture region is formed at the outer periphery of the second mixture region, the vicinity of the combustion chamber wall is an air layer with little fuel,
An internal combustion engine combustion control apparatus characterized in that the injection amount in the second fuel injection is smaller than the first fuel injection amount at least near the knock limit. 2. The combustion control apparatus for an internal combustion engine according to claim 1 , wherein the injection amount in the second fuel injection is set to about 40% or less of the total injection amount at least near the knock limit . The difference in excess air ratio between the first air-fuel mixture region and the second air-fuel mixture region is about 1.0 or less at least near the knock limit. Combustion control device for an internal combustion engine. The internal combustion engine according to any one of claims 1 to 3, wherein the second fuel injection timing is controlled so that the ignition timing of the first air-fuel mixture region comes after the end of the compression stroke. Engine combustion control device. 2. The difference in excess air ratio is reduced, the first fuel injection timing is retarded, or the injection amount in the second fuel injection is decreased in accordance with a decrease in engine load. The combustion control apparatus for an internal combustion engine according to any one of claims 4 to 4.

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