JP2007247625A - Cylinder direct injection internal-combustion engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a cylinder direct injection internal-combustion engine that suppresses the generation of unburned air/fuel mixture to prevent knocking. <P>SOLUTION: The cylinder direct injection internal-combustion engine includes a fuel injecting means 11 for directly injecting fuel into a cylinder; a supply rate controlling means 13 for controlling variably the supply rate of an air/fuel mixture that the fuel injected from the fuel injecting means 11 forms by taking in an ambient air; a combustion initiating means 12 for initiating combustion of the air/fuel mixture during the period of fuel injection; an operating condition detecting means 14, 15 for detecting operating condition of the engine; and a burning rate detecting means 13 for detecting burning rate of the fuel/air mixture depending on the operating condition, and has a combustion mode that forms burning flame of the fuel/air constantly within a combustion chamber 1 by controlling supply rate so as to be substantially equivalent to the burning rate by the supply rate controlling means 13. <P>COPYRIGHT: (C)2007,JPO&INPIT

Description

  The present invention relates to control of combustion in a direct injection type internal combustion engine, and more particularly to control for improving output while preventing occurrence of knocking.

  In a gasoline engine, in the process in which the flame ignited by the spark plug propagates from the vicinity of the spark plug toward the outer wall of the combustion chamber, the unburned mixture mainly in the vicinity of the outer periphery of the combustion chamber is compressed by the piston rise and the propagating flame There is known a knocking phenomenon in which an ignition temperature is reached by a compression action (an increase in combustion chamber pressure due to an explosion), and a self-ignition causes a rapid pressure increase.

  In order to prevent the occurrence of the knocking phenomenon, the upper limit of the compression ratio is limited to such an extent that the unburned mixture does not self-ignite, and as a result, the maximum torque is limited.

  By the way, in an in-cylinder direct injection engine, if the fuel after fuel injection can be ignited and burned quickly before it diffuses, there will be no air-fuel mixture near the outer periphery of the combustion chamber, so knocking will theoretically occur. I can't. Such a combustion mode is generally called spray combustion, and is common in diesel engines.

  However, in the case of a general diesel engine, the position at which the fuel spray starts to burn varies depending on the operating conditions such as the engine load and the engine speed and the EGR amount. For this reason, depending on the operating conditions, smoke and CO may be generated by burning before the fuel is sufficiently mixed with air, and exhaust performance may be deteriorated.

  In addition, as spray combustion targeting a gasoline engine, after generating a swirl flow in the cylinder, the fuel injected from the fuel injection valve ignites immediately after taking in the air in the combustion chamber to form an air-fuel mixture, In addition, Non-Patent Document 1 discloses a technique for localizing a position where combustion starts (hereinafter referred to as a flame surface) at a desired place by balancing the flame propagation speed and the fuel supply speed.

According to Non-Patent Document 1, combustion without generating smoke can be realized by controlling the air-fuel mixture concentration on the flame surface so that the average equivalent ratio of the spray cross section is about 2.
SAE Paper No. 610012

  However, Non-Patent Document 1 does not describe application to an engine in which the atmosphere in the combustion chamber changes every moment like an automobile engine. In other words, the pressure in the cylinder always fluctuates due to the vertical movement of the piston, and the in-cylinder atmosphere changes due to changes in the in-cylinder pressure and the in-cylinder temperature accompanying changes in the engine load and engine speed. There is no description of a technique for localizing the flame to the desired position when the changes.

  Therefore, the standing position of the flame is not always within the cylinder bore diameter, and the theoretical standing position of the flame surface may be outside the cylinder bore. In this case, since the fuel supply speed is higher than the flame propagation speed, even if a flame is generated by spark ignition, the flame surface moves in the direction of the cylinder wall surface, and then the fuel that is injected thereafter reaches the cylinder wall surface before starting combustion. There is a risk of misfire by collision.

  Further, there is no description of control for maintaining the air-fuel ratio at the ignition position at a desired value when the standing position of the flame surface changes with the change in the in-cylinder atmosphere.

  Therefore, as in the case of the diesel engine described above, the rich air-fuel mixture that is not sufficiently mixed with air may burn and produce smoke and CO.

  Therefore, in the present invention, by realizing spray combustion in which a flame surface is fixed in a cylinder bore in a gasoline engine for a vehicle, the generation of knocking is avoided and the output is improved, and the generation of smoke and CO is prevented. The purpose is to improve the exhaust performance.

  The direct injection type internal combustion engine of the present invention has a fuel injection means for directly injecting fuel into the cylinder of the engine, and a supply speed of an air-fuel mixture formed by the fuel injected from the fuel injection means taking in ambient air Supply speed control means for variably controlling, combustion starting means for starting combustion of the air-fuel mixture during the fuel injection period, operating state detecting means for detecting the operating state of the engine, and the mixture of the air-fuel mixture corresponding to the operating state A combustion speed detecting means for detecting a combustion speed, and the supply speed control means controls the supply speed so as to be substantially equal to the combustion speed, whereby the combustion flame of the air-fuel mixture is brought into the cylinder. It has the 1st combustion mode to make it settle.

  According to the present invention, the combustion flame can be localized in the cylinder by controlling the supply speed of the air-fuel mixture to be equal to the combustion speed. If the combustion flame is fixed in the cylinder, the injected fuel starts combustion when it reaches the combustion flame, and combustion in which the fuel injection period is substantially equal to the combustion period is established.

  That is, since no unburned air-fuel mixture is generated, knocking during high load operation can be avoided.

  Thereby, an engine compression ratio can be raised and the fuel consumption performance can be improved.

  In addition, since knocking does not occur, the combustion characteristics do not depend on the self-ignitability of the fuel, and it is sufficient that ignition or ignition can be performed. Therefore, the fuel used is not limited by characteristics such as octane number and cetane number.

  Hereinafter, embodiments of the present invention will be described with reference to the drawings.

  FIG. 1 is a schematic diagram of a system configuration of the present embodiment, and is a cross-sectional view of the periphery of an engine cylinder viewed from the front side of the engine.

  2 is a cylinder head, 3 is a cylinder block, 4 is a piston that slides in a cylinder provided in the cylinder block 3, and 1 is a combustion chamber formed by the lower surface of the cylinder head 2, the cylinder block 3, and the crown surface of the piston 4. . Reference numerals 5 and 6 respectively denote an intake passage and an exhaust passage having an opening in the combustion chamber 1, 7 an intake valve for opening and closing the opening of the intake passage 5, and 8 an exhaust valve for opening and closing the opening of the exhaust passage 6. The intake valve 7 and the exhaust valve 8 are driven by an intake camshaft 9 and an exhaust camshaft 10, respectively. 11 is a fuel injection valve (fuel injection means) for directly injecting fuel into the combustion chamber 1, 12 is an ignition plug (combustion start means) for igniting the air-fuel mixture in the combustion chamber 1, and 13 is the injection timing of the fuel injection valve 11 A control unit (supply speed control means, combustion speed detection means) for controlling the injection amount and the ignition timing of the spark plug 12.

  The fuel injection valve 11 is a so-called multi-hole injection valve having a plurality of injection holes. In addition, the kind of injection valve may be an outward opening injection valve, a swirl injection valve, etc. other than the general injection valve which opens and closes an injection hole with a needle valve.

  The arrangement of the spark plug 12 will be described with reference to FIGS. 2 is a top view of FIG. 1, in which the intake valve 7, the exhaust valve 8, the intake camshaft 9, and the exhaust camshaft 10 are omitted.

  The spark plugs 12 are provided in the same number as the injection holes so that each of the plurality of fuel sprays injected from the fuel injection valve 11 can be ignited. For example, as shown in FIGS. 1 and 2, it is provided at a position that does not interfere with the intake passage 5, the exhaust passage 6, and the fuel injection valve 11 and that can ignite the fuel spray.

  In the present embodiment, since the fuel injection directions are six as shown in FIG. 2, six spark plugs 12 are provided at positions downstream of the fuel injection direction as viewed from the fuel injection valve 11.

  The fuel injection direction, the number of fuel sprays, and the arrangement of the spark plugs 12 shown in FIGS. 1 and 2 are merely examples.

  Next, the outline of the combustion mode applied in this embodiment will be described with reference to FIG.

  The upper diagram of FIG. 3 is a diagram schematically showing the behavior from when the fuel is injected to ignition and combustion, and the lower diagram of FIG. 3 shows the relationship between the distance from the fuel injection hole and the speed of the fuel spray. It is. The fuel spray speed is expressed as an average flow velocity.

  As shown in the lower diagram of FIG. 3, the spray speed has a characteristic of decreasing as the distance from the fuel injection hole increases, and there is a point P that balances with the flame propagation speed in the middle of the decrease. The point P differs depending on the atmosphere in the combustion chamber 1, the fuel injection pressure, and the like.

  The fuel injected from the fuel injection valve 11 forms an inflammable mixture through atomization and evaporation while taking in ambient air. When a spark is ignited on the combustible air-fuel mixture by the spark plug 12 provided at the point Q downstream of the fuel spray, a flame is generated.

  As shown in the lower diagram of FIG. 3, since the spray speed is lower than the flame propagation speed at the point Q, the flame propagates also in the direction of the fuel injection hole while continuing the fuel injection. Then, the propagation of the flame in the direction of the fuel injection hole stops at the point P where the spray speed and the flame propagation speed balance, and thereafter, the injected fuel starts to burn at the point P. Assuming that premixed combustion continues for a very short time after the start of combustion, the reaction zone of premixed combustion becomes clear. Therefore, when this reaction zone is defined as a flame surface, the flame surface is fixed at a point P until the fuel injected just before the end of fuel injection burns.

  In the combustion of the present embodiment, the form of the combustion flame is not particularly limited as long as combustion with good exhaust performance is obtained without generating an unburned air-fuel mixture, so the flame surface does not necessarily need to be clearly defined.

  Further, spark ignition may be performed once after the start of fuel injection, as in a normal gasoline engine.

  As described above, the injected fuel starts to burn immediately after forming an air-fuel mixture, and since all of the injected fuel starts to burn on the flame surface, spray combustion that does not generate unburned air-fuel mixture becomes possible, Further, the fuel injection time and the combustion time are almost equal.

  Hereinafter, such combustion is referred to as “mixed gas rule combustion” (first combustion mode). Note that, in the case of gas mixture law combustion, fuel injection is performed at a time when the crankshaft angle is close to the compression top dead center.

  According to the air-fuel mixture combustion, since the unburned mixture is not generated as described above, the unburned mixture existing in the vicinity of the cylinder periphery is self-ignited before the combustion by the flame propagation, and the knocking phenomenon occurs. The problem that can not occur. That is, knocking does not occur as long as the gas mixture combustion is established.

  In addition, when the spark ignition position by the spark plug 12 is on the nozzle hole side from the point P, the spray velocity is higher than the flame propagation velocity, so the flame propagates only in the direction away from the nozzle hole and ignites at the point Q. As in the case, it is fixed at a point P where the flame propagation speed and the fuel supply speed are balanced.

  Thus, even if the ignition position by the spark plug 12 is different from the place where the flame surface is fixed, the flame surface is fixed at a position where the spray speed and the flame propagation speed are balanced. Need only be a position where the fuel spray can be ignited by sparks, and the distance from the fuel injection valve 11 may not be set strictly.

  Further, the flame propagation speed, that is, the combustion speed of the injected fuel, has a characteristic that varies depending on the in-cylinder temperature, the in-cylinder pressure, the in-cylinder gas flow, and the air-fuel mixture concentration. Specifically, there is a relationship as shown in FIG.

  FIG. 4A shows the relationship between the combustion speed and the in-cylinder temperature, with the vertical axis representing the combustion speed and the horizontal axis representing the in-cylinder temperature. As shown in the figure, there is a characteristic that the combustion speed increases as the in-cylinder temperature rises.

  FIG. 4B is a diagram showing the relationship between the combustion speed and the in-cylinder pressure, where the vertical axis represents the combustion speed and the horizontal axis represents the in-cylinder pressure. As shown in the figure, there is a characteristic that the combustion speed decreases as the in-cylinder pressure increases.

  FIG. 4C shows the relationship between the combustion speed and the in-cylinder gas flow, where the vertical axis represents the combustion speed and the horizontal axis represents the flow strength. As shown in the figure, there is a characteristic that the combustion speed increases as the flow becomes stronger. The gas flow mentioned here includes not only swirl flow and tumble flow that are intentionally generated by providing a control valve, but also flow generated due to the shape of the intake port, the valve lift amount, and the like.

  FIG. 4D is a diagram showing the relationship between the combustion speed and the fuel injection pressure, with the vertical axis representing the combustion speed and the horizontal axis representing the fuel injection pressure. As the fuel injection pressure increases, the amount of fuel injected during the same injection time increases and the mixture concentration becomes richer, so the combustion speed increases. That is, the combustion speed changes with the change of the fuel injection pressure, and the position where the flame surface stays changes. However, in the gas mixture law combustion, the flame surface does not always have to be fixed at a certain position during the combustion period, and it is only required to be in a range satisfying a condition described later.

  Here, the position where the flame surface is fixed will be described with reference to FIG.

  In FIG. 9, the vertical axis represents the average spray flow velocity Ux and the horizontal axis represents the spray distance. The solid lines High and Plow represent the characteristics at high fuel pressure and low fuel pressure, respectively. Dotted lines Stigh and Stlow represent flame propagation speeds St at high fuel pressure and low fuel pressure, respectively. That is, a point Lb2 at which the average flow rate and the combustion speed (flame propagation speed) are balanced at high fuel pressure, and a point Lb1 at which the average flow rate and the combustion speed are balanced at low fuel pressure are respectively at high fuel pressure. Flame standing distance at low fuel pressure.

  The average spray flow velocity Ux at the same spray distance is naturally higher at the high fuel pressure than at the low fuel pressure. Further, as shown in FIG. 4D, the flame propagation speed St has a characteristic that it increases as the fuel pressure increases. Therefore, as shown in FIG. 9, the flame standing distance Lb becomes larger at the high fuel pressure than at the low fuel pressure.

  By the way, the generation of smoke can be cited as a problem similar to the generation of unburned air-fuel mixture during combustion. The amount of smoke produced is related to the fuel equivalent ratio φ as shown in FIG.

  FIG. 7 shows the equivalence ratio on the horizontal axis and the amount of smoke generated on the vertical axis. As apparent from FIG. 7, when the equivalent ratio φ is approximately 2 or less, smoke is hardly generated.

  Therefore, in the case of gas mixture law combustion, if the average equivalence ratio φx of the spray cross section of the flame surface is approximately 2 or less, the occurrence of smoke can be prevented.

  The relationship between the average equivalent ratio φx of the spray cross section and the distance from the fuel injection hole (hereinafter referred to as the spray distance) will be described with reference to the lower diagram of FIG.

  The lower graph of FIG. 6 shows the characteristics (θ1, θ2, θ3 in the figure) when the vertical axis represents the equivalent ratio φx of the spray cross section, the horizontal axis represents the spray distance, and the spray spread angle θ is different. Note that θ2 and θ3 represent cases where the spray angle θ is twice and three times θ1, respectively, and θ1, θ2, and θ3 are cases where the same fuel amount is injected.

  As shown in the lower diagram of FIG. 6, the equivalence ratio decreases as the spray distance increases. This is because ambient air is taken in as the spray distance increases. When the same amount of fuel is injected, the fuel spray volume increases as the spray spread angle θ increases, so that mixing with the surrounding air is promoted, and the equivalence ratio φ is approximately at a position close to the injection hole. 2 or less.

  Therefore, if the smoke limit distances for the spray angles θ1, θ2, and θ3 are Ls1, Ls2, and Ls3, respectively, the relationship of Ls1> Ls2> Ls3 is established.

  The spray spread angle θ is basically determined by the opening direction of the injection hole of the fuel injection valve 11, but as shown in the upper diagram of FIG. 6, the spread angle θ tends to increase as the atmosphere density increases. is there. Therefore, the higher the atmosphere density, the shorter the spray distance at which the equivalent ratio φx = 2.

  From the above, it can be seen that smoke generation at the time of combustion can be prevented if the average equivalent ratio φx on the flame surface is 2 or less. However, since the combustibility deteriorates as the equivalent ratio φ decreases, the equivalent ratio that can be combusted has a lower limit (lean flammability limit).

  Therefore, the range from φx = 2 to the lean combustible limit is a range in which combustion without generating smoke is possible (smokeless combustion range).

  FIG. 8 shows the relationship between the average equivalent ratio φx of the spray cross section, the spray distance, and the smokeless combustion range. In FIG. 8, the vertical axis represents the spray average equivalent ratio φx, the horizontal axis represents the spray distance, and the smokeless combustion range is represented by the hatched portion.

  In FIG. 8, when the spray distance is between SL1 and SL2, the average equivalent ratio φx of the spray cross section falls within the smokeless combustion range. The spray distance at which the average equivalence ratio φx of the spray cross section is 2 or less and the lean combustible limit or more is set as the establishment distance of the smokeless combustion as between SL1 and SL2. The spray distance of the portion closest to the fuel injection hole in the smokeless combustion range, that is, the portion where the average equivalent ratio φx of the spray cross section is 2 is defined as the smoke limit distance Ls.

  From the above, if the flame standing distance Lb is within the smokeless combustion establishment distance, the combustion does not cause knocking and does not generate smoke.

  However, in the case of a vehicle engine, since the operating conditions such as the engine load and the engine speed change every moment, the flame standing position also changes every moment. Also, since the bore size of the cylinder is fixed, the flame surface must be fixed in the cylinder as will be described later.

  Therefore, the control described below is performed in order to perform the gas mixture law combustion in the vehicle engine.

  In addition, the size of the injection hole of the fuel injection valve 11 is such that the average equivalence ratio φx of the spray cross section is approximately 2 or less at any position in the combustion chamber by the control for performing the gas mixture law combustion described later. Set so that a mixture is formed. For example, the variation range of the in-cylinder atmosphere due to changes in engine load, engine speed, etc. is obtained in advance, and the average equivalence ratio φx is set to a value that can be approximately 2 or less in the widest possible range. .

  FIG. 5 is a flowchart showing a control routine executed by the control unit 13 to perform the gas mixture combustion. In this control, homogeneous combustion is performed when the flame surface standing position does not fit in the cylinder bore even by adjustment described later. Homogeneous combustion is a combustion mode in which a fuel amount determined according to the operating state is injected during the intake stroke, diffused and mixed (premixed) with the air in the cylinder during the compression stroke, and spark-ignited near the compression top dead center ( (Second combustion mode).

  Note that this control routine is performed in an engine low speed / high load operation region where knocking is likely to occur in the premixed homogeneous combustion, for example, a region surrounded by hatching in the operation region map of FIG. Homogeneous combustion is performed in other areas (including high-speed and high-load operation areas).

  In step S101, the in-cylinder temperature, pressure, and atmosphere density are read. Specifically, a map in which each value is assigned to the engine speed and the load is created, and the speed and load detected by the speed sensor 14, the accelerator opening sensor 15 and the like (operating state detecting means) are used. Search the map. The in-cylinder temperature, pressure, and atmosphere density refer to the temperature, pressure, and atmosphere density at the time of compression top dead center, respectively.

  In step S102, the fuel spray spread angle θ is calculated based on the atmosphere density read in step S101, and the smoke limit distance Ls is calculated based on the spray spread angle θ and the nozzle hole diameter of the fuel injection valve.

  Between the atmosphere density and the spray spread angle θ, as shown in the upper diagram of FIG. 6, there is a characteristic that the spray spread angle increases as the atmosphere density increases, that is, as the actual compression ratio in the cylinder during fuel injection increases. is there. Therefore, a map as shown in the upper diagram of FIG. 6 is created in advance based on the specifications of the fuel injection valve 11, and the fuel spray spread angle θ is obtained by searching with the atmosphere density read in step S101.

  In step S103, it is determined whether or not the smoke limit distance Ls is smaller than the cylinder bore radius. As a result of the determination, if the smoke limit distance is larger, the process proceeds to step S115, where it is determined that homogeneous combustion is performed in the same manner as a general gasoline engine, and the process is terminated.

  The cylinder bore radius here refers to the distance from the fuel injection hole to the cylinder wall surface.

  If the smoke limit distance Ls is smaller as a result of the determination in step S103, the process proceeds to step S104, and initial setting of the fuel injection pressure (fuel pressure) is performed. Specifically, the minimum fuel pressure determined from the specifications of an engine system such as a fuel pump (not shown) is set.

  In step S105, an average flow velocity Ux of the spray cross section is obtained.

  Specifically, the relationship between the fuel pressure and the average flow velocity Ux at the time of injection at that fuel pressure is obtained in advance and mapped, and the map is searched with the set fuel pressure. In addition, you may calculate by the following Formula (1).

  In step S106, the flame propagation speed St, that is, the combustion speed of the injected fuel is obtained. For example, the relationship between the combustion speed and the fuel injection pressure (fuel pressure) as shown in FIG. 4 (d) is mapped and can be obtained by searching with the set fuel pressure.

  In step S107, the flame standing distance Lb, that is, the distance from the fuel injection hole to the position where the flame propagation speed St and the average fuel flow velocity Ux are balanced is obtained.

  The spray average flow velocity Ux at the same spray distance is naturally higher at the high fuel pressure than at the low fuel pressure, and the flame propagation speed St is higher as the fuel pressure is higher as shown in FIG. is there. Therefore, as shown in FIG. 9, the flame standing distance Lb becomes larger at the high fuel pressure than at the low fuel pressure.

  In step S108, it is determined whether the flame standing distance Lb calculated in step S107 satisfies the relationship of the following expression (2).

Lb = bore radius × 0.9 (2)
The reason why 0.9 is multiplied in the formula (2) is as follows.

  In order to perform gas mixture law combustion, it is a necessary condition that the flame standing distance Lb is equal to or less than the bore radius of the cylinder. Therefore, even when the flame standing distance Lb is the same as the bore radius, it is theoretically possible to perform air-fuel mixture combustion. However, the fact that the flame standing distance Lb is the same as the bore radius means that the flame surface is formed on the cylinder wall surface. When combustion is performed in such a state, a cooling loss from the cylinder wall surface increases and there is a risk of misfire.

  In other words, it is practically possible to execute the mixture combustion in the case where the position where the flame surface is formed is separated from the cylinder wall by an appropriate distance. In this embodiment, the appropriate distance is set to the bore. It is set as about 10% of the radius.

  Therefore, the numerical value to be multiplied by the equation (2) is not limited to 0.9, and may be a value that allows the flame standing distance Lb to be an upper limit value in which the cooling loss on the cylinder wall surface is allowable.

  If the result of the determination in step S108 is affirmative, the process proceeds to step S109, where it is determined to perform mixture gas mixture combustion, and the process ends. If the determination result is negative, the process proceeds to step S110 to determine whether or not the relationship of the following equation (3) is satisfied.

Lb <bore radius × 0.9 (3)
As a result of the determination, if the flame standing distance Lb is smaller than 0.9 times the bore radius, the process proceeds to step S111, and if larger, the process proceeds to step S113.

  In step S111, it is determined whether or not the current fuel pressure is an upper limit value determined from the specifications of the engine system.

  When it determines with it being an upper limit, it progresses to step S109 and determines performing mixture gas mixture combustion, and complete | finishes a process.

  When it determines with it not being an upper limit, it progresses to step S112, raises a fuel pressure, and returns to step S105.

  By performing the steps S111 and S112, feedback control is performed so that the maximum fuel pressure is obtained in a range satisfying the expression (2). At this time, the fuel injection amount per time increases by the amount corresponding to the increase in the fuel pressure, so that the fuel injection injection pulse is shortened in accordance with the increase in the fuel pressure in order to inject the fuel injection amount according to the operating state. That is, the injection pulse is the shortest within a range satisfying the expression (2). Thereby, both output improvement and exhaust performance can be achieved.

  If the result of determination in step S110 is negative, the process proceeds to step S113, where it is determined whether or not the fuel pressure is a lower limit value determined from the specifications of the engine system. When it determines with it being a lower limit, it progresses to step S115, determines performing homogeneous combustion, and complete | finishes a process. If it is determined that the fuel pressure is not the lower limit value, the process proceeds to step S114, the fuel pressure is decreased, and the process proceeds to step S105.

  When the fuel pressure is lowered, the flame standing distance Lb approaches the fuel injection hole as shown in FIG. 9, so that by performing the above steps S113 and S114, the fuel pressure becomes a lower limit value determined from the specifications of the engine system. The feedback control is performed so that the formula (2) is established within the range. When the formula (2) is established in step S108 by this feedback control, the maximum fuel pressure that satisfies the formula (2) is obtained.

  The above control is summarized as follows.

  By controlling the supply speed (fuel pressure) of the air-fuel mixture in accordance with the combustion speed (flame propagation speed) that changes according to the operating state of the engine, the flame surface can be fixed in the combustion chamber over a wide operating range.

  By setting the fuel pressure to the maximum value within the range where the flame surface is fixed in the combustion chamber, the combustion period can be shortened as much as possible, and both the engine output and the exhaust performance can be achieved.

  The injected fuel begins to burn as soon as it forms an air-fuel mixture, and all the fuel injected after the flame front is formed starts to burn on the flame front, so all of the fuel injected during the stroke burns. As a result, no unburned mixture is generated.

  By forming the flame surface away from the cylinder wall, it prevents the fuel from colliding with the cylinder wall, thus avoiding the generation of wall flow on the cylinder wall surface and the deterioration of exhaust performance due to the flame extinguishing near the cylinder wall surface. can do.

  By the way, as described above, knocking does not occur during the execution of gas mixture law combustion, so that the compression ratio can be increased to a value at which knocking occurs in homogeneous combustion. Therefore, a variable compression ratio mechanism (actual compression ratio variable means) is used to achieve a higher compression ratio than ordinary homogeneous combustion during mixed gas rule combustion, and compression equivalent to ordinary homogeneous combustion during homogeneous combustion. It is also possible to control the ratio.

  As a variable compression ratio mechanism, for example, a mechanism as shown in FIG. 11 is known. Since FIG. 11 is a mechanism described in Japanese Patent Application Laid-Open Nos. 2001-227367 and 2002-61501, only the outline of the mechanism will be described.

  In this mechanism, the piston 4 is connected to the crankshaft 23 via the first link 20 and the second link 21. The first link 20 and the second link 21 are connected via a connecting pin 26. The center of the second link 21 is rotatably connected to the crankpin 28 of the crankshaft 23 and rotates together with the crankshaft 23. A third link 22 is rotatably connected to the second link 21 on the opposite side of the first link 20 via a connecting pin 27, and the third link 22 is fixed to the control shaft 24 via a connecting pin 29. Is done. The center axis of the control shaft 24 and the fastening portion of the third link 22 are eccentric, and the rotation of the control shaft 24 moves the connecting pin 29 and changes the inclination of the second link 21 to change the first link. 20 and the top dead center position of the piston 4 are changed. The control shaft 24 is rotated by an actuator 25 with a motor.

  The change in the top dead center position due to the rotation of the control shaft 24 will be described with reference to FIG. FIG. 12 is a diagram schematically showing the positional relationship among the links, the connecting pins, and the control shaft 24. The left figure of FIG. 12 shows a state where the top dead center position of the piston 4 is high, that is, set to a high compression ratio, and the right figure shows a state where the top dead center position of the piston 4 is low, ie, set to a low compression ratio.

  When the connecting pin 29 moves in a direction lowering with respect to the central axis of the control shaft 24 by rotating the control shaft 24, the position of the connecting pin 27 is also lowered, and the second link 21 is centered on the crank pin 28 in the figure. Tilt clockwise. As a result, the position of the connecting pin 26 rises, the piston 4 also rises, and the compression ratio increases.

  Conversely, when the connecting pin 29 moves in a direction higher than the central axis of the control shaft 24, the position of the connecting pin 27 also rises, and the second link 21 is tilted counterclockwise in the drawing with the crank pin 28 as the center. As a result, the position of the connecting pin 26 is lowered, the piston 4 is also lowered, and the compression ratio is lowered.

  Normally, in the mechanism described above, the engine is set to a low compression ratio state to prevent knocking regardless of the engine speed when the engine is in a high load operation region, and in the low and medium load operation region where the possibility of knocking is low, In order to improve, the high compression ratio state is set.

  However, according to the present embodiment, even if it is a high load operation region, if it is a region where the engine speed is low, the mixture gas rule combustion is executed, so knocking does not occur. Therefore, even in the high load operation region, if the engine speed is low, it can be set to a high compression ratio state, and higher output can be achieved.

As described above, according to the present embodiment, the following effects can be obtained.
(1) Since the supply speed is controlled so as to be substantially equal to the combustion speed (flame propagation speed), the combustion flame of the air-fuel mixture is controlled to be localized in the cylinder bore, so the fuel injection period and the combustion period When combustion is substantially equal, and the injected fuel forms an air-fuel mixture, it immediately burns. Thereby, since generation | occurrence | production of unburned air-fuel | gaseous mixture can be suppressed, generation | occurrence | production of knocking at the time of high load operation | movement can be avoided. In particular, although the required output is large, it is possible to improve the output at low engine speed and high load, which is limited to prevent knocking. Furthermore, by suppressing the occurrence of knocking, it is not necessary to consider self-ignitability such as octane number and cetane number when selecting fuel. Further, since the ignition timing is set during the fuel injection period, it is possible to suppress the diffusion of the air-fuel mixture as much as possible, thereby suppressing the unburned air-fuel mixture from staying around the cylinder wall that is the cause of knocking. . Moreover, since generation | occurrence | production of unburned air-fuel mixture can be suppressed, exhaust performance and fuel consumption performance can be improved.
(2) Since the supply speed of the air-fuel mixture is controlled by adjusting the fuel pressure, control with good responsiveness according to the engine operating state is possible. The combustion period, that is, the engine load is controlled by controlling the injection period and the injection rate.
(3) Since the supply speed of the air-fuel mixture is controlled so that the position where the flame surface is fixed is located downstream of the fuel spray from the position where the fuel is generally in the gas phase, the fuel is generally oxygen (air in the combustion chamber). ), The combustion is performed in a state where the mixing is completed, and combustion with good combustion efficiency and good exhaust performance can be realized.
(4) Since the supply rate of the air-fuel mixture is controlled so that the average equivalence ratio φx of the spray cross section at the position where the flame surface is fixed is approximately 2 or less, combustion without generating smoke or CO is possible.
(5) Since the highest fuel pressure is set within the range that satisfies the condition that the flame surface is settled in the cylinder bore, the combustion period can be shortened and the engine operation with high fuel efficiency and high output can be achieved. Is possible. In addition, fuel injection at high fuel pressure has the effect of atomizing the fuel spray, promoting the formation time of the air-fuel mixture, and increasing the turbulence intensity, so the combustion speed is improved and smoke is easily generated. An effect of suppressing the formation of smoke by promoting the formation of the air-fuel mixture during operation is also obtained.
(6) Since the injected fuel is spark-ignited, the combustion start timing can be set to a desired timing regardless of the operating conditions.
(7) Since the fuel injection valve 11 is a multi-hole type injection valve, each injection hole can be set small and injection can be performed at a high pressure, fuel spray atomization, promotion of the mixture formation timing, and even disturbance. The effect of increasing the flow strength can be obtained.
(8) By setting the nozzle hole diameter of the fuel injection valve 11 such that the average equivalence ratio φx of the spray cross section is approximately 2 or less, combustion without generating smoke or CO becomes possible.
(9) Since gas mixture combustion is performed at least in the low rotation / high load operation region, it is possible to improve the output performance in the operation region where the output is conventionally limited to prevent knocking.
(10) A variable compression ratio mechanism is provided, and the actual compression ratio of the engine is set higher than that in premixed combustion in which fuel is injected in the intake stroke and ignited in the vicinity of the compression top dead center at the time of execution of air-fuel mixture combustion. As a result, it becomes possible to drive with high thermal efficiency while suppressing the occurrence of knocking, and to improve fuel efficiency.

  A second embodiment will be described.

  The configuration of the engine system of this embodiment is the same as that of the first embodiment. In addition, in order to perform gas mixture law combustion, a method of performing feedback control of the condition that the flame surface is settled in the cylinder bore by the smoke limit distance Lb, the spray average flow velocity Ux, and the flame propagation velocity St is the same as in the first embodiment. It is. However, in the first embodiment, the fuel pressure is controlled to be as large as possible regardless of the engine load in a range where combustion without generating smoke is possible, whereas in this embodiment, the injection is performed regardless of the load. The difference is that the fuel pressure is controlled so that the period is constant.

  Hereinafter, specific control will be described with reference to FIG.

  FIG. 13 is a flowchart showing a control routine executed by the control unit 13 for the gas mixture law combustion.

  Steps S201 to S203 are the same as steps S101 to S103 in FIG.

  In step S204, the fuel pressure is set according to the engine load. Specifically, a map in which the fuel pressure is assigned to the engine load is created, and the map is searched using the engine load detected by the rotational speed sensor 14 or the accelerator opening sensor 15 or the like.

  Here, a map in which the fuel pressure is allocated so that the fuel injection period is constant regardless of the engine load is used. The fixed value of the target fuel injection period will be described later.

  Specifically, as shown in FIG. 14, the fuel pressure increases as the engine load increases. The fuel pressure upper limit value in FIG. 14 is the upper limit value of the fuel pressure determined by the specifications of the engine system. Therefore, when the fuel pressure necessary for making the fuel injection period constant exceeds the upper limit value, the fuel pressure period is fixed to the upper limit value.

  If the fuel pressure is set in step S204, the process proceeds to step S205, but steps S205 to S210 are the same as steps S105 to S110 in FIG.

  If the determination result in step S210 is affirmative, the process proceeds to step S209, where it is determined that the mixture gas mixture combustion is performed, and the process ends.

  If the determination result in step S210 is negative, the process proceeds to steps S211 and S212, and the same processing as steps S113 and S114 in FIG. 5 is performed.

  By the above control, the fuel injection period is basically constant regardless of the engine load, unless the upper limit on the system is exceeded.

  If the fuel pressure is constant regardless of the engine load, there are the following problems.

  As the engine load increases, the amount of fuel injection required also increases, so the fuel injection period increases as the engine load increases, and in the case of mixture gas combustion, the fuel injection period and the combustion period It is almost equivalent. Therefore, as the engine load increases, the combustion time becomes longer, which is contrary to the output demand for burning the injected fuel as quickly as possible.

  In addition, when the engine load is low, the drive loss ratio of the fuel pump becomes large. Therefore, if fuel injection is performed at a fuel pressure higher than necessary, that is, if the fuel injection period is shorter than necessary, the fuel loss performance due to the drive loss. There is also a risk of worsening. Note that when the engine load is low, the fuel injection amount is small, so even if the fuel pressure is lowered, the combustion period does not increase greatly, and the rebound to fuel efficiency is small.

  Therefore, by controlling the fuel injection period to be a constant value that satisfies both the demand at the time of engine high load operation and the demand at low engine load, the fuel injection period can be extended when the engine load is high. It can be suppressed to satisfy the demand from the output side, and the deterioration of fuel efficiency can be suppressed without increasing the fuel pressure more than necessary.

  The constant value of the fuel injection period that satisfies the above requirements varies depending on the engine system specifications such as the normal rotation speed range, the output, the capacity of the fuel pump, and the like, and therefore, a suitable value is set for each engine system.

  As described above, in this embodiment, the following effects can be obtained in addition to the effects similar to the effects described in (1) to (4) and (6) to (9) of the first embodiment.

  Since the supply speed of the air-fuel mixture is controlled so that the fuel injection period becomes substantially constant even when the engine load changes, it is possible to suppress the deterioration of the equal volume due to the increase in the injection period when the engine load increases. Further, in a low load operation region where the total injection amount of fuel is small and the injection period is short, and the long combustion period is not a problem, the fuel pump drive loss can be reduced and the fuel efficiency performance can be improved.

  A third embodiment will be described.

  The third embodiment is basically the same configuration as the first embodiment, except that only one spark plug 12 is provided near the fuel injection valve 11 and the number of injection holes of the combustion injection valve 11 is the same. It differs in that there are more points than that of the first embodiment.

  The arrangement of the spark plug 12 will be described with reference to FIG. The upper diagram of FIG. 15 is a schematic diagram showing the positional relationship between the fuel injection valve 11 and the spark plug 12, and the lower diagram is an AA arrow view of the upper diagram. The AA section is a section including the plug gap position of the spark plug 12.

  30 is a plug gap position of the spark plug 12, and 31a to 31e are fuel sprays (spray cross sections). As for the cross section of the fuel spray, eight spray cross sections are arranged in a substantially circular shape at equal intervals. This substantially circular shape is a spray outline. And the spark plug 12 is located in the vicinity of any one spraying cross section.

  The combustion mode in the above configuration will be described with reference to FIG.

  FIG. 17 is a time chart showing a drive signal transmitted from the control unit 13 to the fuel injection valve 11 during one stroke. As shown in FIG. 17, multistage injection is performed, t1 to t2 being a small amount of pilot injection, and t3 to t4 being a main injection.

  At the time of pilot injection, the fuel pressure is reduced, and spark ignition is performed by the spark plug 12 at the time t2 when the pilot injection ends.

  In addition, the space | interval of the fuel injection valve 11 and the plug gap of the spark plug 12 should just be a space | interval which can ignite the fuel spray 31a reliably at the time t2 of pilot injection completion. Further, the interval from the pilot injection end time t2 to the main injection start timing t3 is set to be shorter than the period during which fuel injected by pilot injection described later is combusting.

  As described above, when spark ignition is performed at the end of pilot injection t2 to generate a flame in the fuel spray, the fuel supply speed becomes lower than the flame propagation speed by reducing the fuel pressure, so between t1 and t2. The flame propagates to the fuel spray that has already been injected.

  Since the fuel spray 31a diffuses as it proceeds downstream of the spray, the flame propagates from the fuel spray 31a to the adjacent fuel spray 31b, that is, in the circumferential direction of the fuel spray outline. And it propagates similarly from the fuel spray 31b to the fuel sprays 31c, 31d, 31e.

  When the fuel spray of the main injection reaches the place where the generated flame is burning, the flame of the fuel sprays 31a to 31e becomes a fire type, and the fuel injected by the main injection starts to burn, Thereafter, the combustion start position moves to the flame standing position.

  In general, near the fuel injection hole during the fuel injection period, the flow rate of fuel spray is high and spark ignition is difficult. However, as described above, the pilot injection flow rate is lowered to facilitate ignition, and the pilot spray fuel sprays 31a to 31a. By using the flame formed by the combustion of 31e as a fire type, the fuel spray of the main injection can be reliably ignited.

  In this way, if all sprays can be burned by one spark plug 12, the layout restriction of the spark plug 12 in the cylinder head 2 is the same as that of a general engine.

  Therefore, since the valve diameters of the intake valve 7 and the exhaust valve 8 are not limited because a large number of spark plugs 12 are provided, the output can be improved by increasing the valve diameter.

  It is desirable that the number of fuel injection holes of the fuel injection valve 11, that is, the number of fuel sprays be large. This is because it is desirable that the interval between the fuel sprays in the circumferential direction is short considering flame propagation in the circumferential direction of the spray outline.

  As described above, in the present embodiment, in addition to the same effects as those in the first embodiment and the second embodiment, the following effects can be further obtained.

  Since the spark plug 12 is disposed on any one or two of the fuel sprays injected from the fuel injection valve 11 so as to ignite in the vicinity of the fuel injection valve 11, the ignited fuel spray is replaced with another fuel spray. As a result, all fuel sprays can be ignited. Thereby, it is not necessary to provide the spark plug 12 with the number of fuel injection holes (the number of sprays). Further, the layout of the spark plug 12 in the cylinder head 2 is facilitated.

  Note that the plug gap position 30 is not limited to that shown in FIG. 15. As shown in FIG. 16, even when the plug gap position of the spark plug 12 is on the periphery of the spray outline, As with, fuel spray can be reliably ignited.

Further, the ignition device is not limited to the spark plug 12 that performs general spark ignition, and a device capable of more powerful volume ignition and line ignition, such as that seen in laser ignition, can also be used.
A fourth embodiment will be described.

  This embodiment is basically the same configuration as the first embodiment, but the structure of the fuel injection valve 11 is different.

  18 is a cross-sectional view (upper view in FIG. 18) and an external view (lower view in FIG. 18) of the tip of the fuel injection valve 11 used in the present embodiment.

  A fuel chamber 45 is provided in the fuel injection valve 11 and is filled with fuel, 40 is a needle valve that is reciprocated in the axial direction of the fuel injection valve 11 in the fuel chamber 45, and 41 is a fuel chamber 45. A fuel injection hole communicating with the outside of the fuel injection valve, 44 is a cap attached to the tip of the fuel injection valve 11, and 46 is formed between the cap 44 and the tip of the fuel injection valve 11 when the cap 44 is attached. Air passages (air supply means) 42 and 43 are air-fuel mixture injection holes and air introduction holes that connect the air passage 46 and the combustion chamber 1, respectively. Therefore, the internal pressure of the air chamber 46 is substantially equal to the internal pressure of the combustion chamber 1.

  The needle valve 40 reciprocates based on a signal input from the control unit 13, and normally the tip of the needle valve 40 is seated on the tip of the fuel chamber 45 and moves upward in the figure when an injection signal is input. To do. When seated, the needle valve 40 closes the fuel injection hole 41. When an injection signal is input, the fuel injection hole 41 is opened.

  In the air-fuel mixture holes 42, the outline X connecting the approximate centers of the air-fuel mixture holes 42 is closer to the fuel injection valve 11 (upper side in FIG. 18) than the outline Y connecting the approximate centers of the air introduction holes 43. Provide as follows. Further, the air-fuel mixture hole 42 and the air introduction hole 43 both communicate with the air passage 46, and the air-fuel mixture hole 42 faces the fuel hole 41 when the cap 44 is attached to the fuel injection valve 11. Positioned to be in position.

  In such a configuration, when the needle valve 40 moves and the fuel injection hole 41 opens, the fuel is injected from the fuel injection hole 41 through the air chamber 46 and the air-fuel mixture injection hole 42. At this time, due to the ejector effect caused by the ejection of fuel, the air in the air chamber 46 is also ejected from the air-fuel mixture nozzle hole 42, and the fuel-air mixture is injected from the air-fuel mixture hole 42. Since air is introduced from the combustion chamber 1 into the air chamber 46 through the air introduction hole 43, the air-fuel mixture is always injected during fuel injection.

  For this reason, it becomes possible to form a uniform combustible air-fuel mixture quickly and to reduce the smoke limit distance Ls as compared with the case of fuel injection with a normal multi-hole injection valve. That is, it is possible to make the flame surface settle at a position closer to the fuel injection valve 11.

  As a result, the supply speed of the air-fuel mixture, that is, the fuel pressure can be lowered without deteriorating the smoke performance, so that the drive loss of the fuel pump can be reduced, and the fuel efficiency and exhaust performance can be improved. .

  It should be noted that an air supply hole is provided in the vicinity of the fuel injection hole of the fuel injection valve 11, and a part of the intake air flowing through the intake passage is supplied to the fuel spray injected from the fuel injection hole (for example, a so-called air assist injection valve (for example, The same effect can be obtained by using JP 2000-213439 A).

  As described above, in the present embodiment, in addition to the same effects as those of the first to third embodiments, the following effects can be further obtained.

  The fuel injection valve 11 has an air passage 46 and supplies air to the fuel injected from the fuel injection hole 41, so that fuel and air are injected from the mixture injection hole 42 in a mixed state. As a result, a uniform combustible air-fuel mixture can be formed more quickly than a fuel injection valve that does not have an air passage 46.

  The present invention is not limited to the above-described embodiments, and it goes without saying that various modifications can be made within the scope of the technical idea described in the claims.

It is the schematic of the system configuration | structure of this embodiment. It is a combustion chamber top view for demonstrating a combustion form. It is a figure showing the relationship between spray speed and nozzle hole distance. (A)-(d) is a figure showing the relationship between a combustion speed and cylinder interior. It is a flowchart showing the control routine of 1st Embodiment. The upper diagram shows the relationship between the spray spread angle and the atmosphere density, and the lower diagram shows the relationship between the average equivalence ratio and the spray distance. It is a figure showing the relationship between an average equivalent ratio and the amount of smoke generation. It represents the relationship between the smokeless combustion range and the average equivalent ratio. It is a figure showing the relationship between the average flow velocity of fuel spray, and spray distance. It is a map which shows the operation area | region which performs air-fuel | gaseous law combustion. It is a figure showing an example of a variable compression ratio mechanism. It is the schematic of the compression ratio change operation | movement of a variable compression ratio mechanism. It is a flowchart showing the control routine of 2nd Embodiment. It is the map which allocated fuel pressure to engine load. It is a figure showing the position of the spark plug of 3rd Embodiment (the 1). It is a figure showing the position of the spark plug of 3rd Embodiment (the 2). It is a figure showing the fuel-injection pulse and fuel pressure of 3rd Embodiment. It is the schematic of the structure of the fuel injection valve of 4th Embodiment.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Combustion chamber 2 Cylinder head 3 Cylinder block 4 Piston 5 Intake passage 6 Exhaust passage 7 Intake valve 8 Exhaust valve 9 Intake camshaft 10 Exhaust camshaft 11 Fuel injection valve 12 Spark plug 13 Control unit 20 First link 21 Second link 22 Third link 23 Crankshaft 24 Control shaft 25 Actuator 40 Needle valve 41 Fuel injection hole 42 Air mixture injection hole 43 Air introduction hole 44 Cap 45 Fuel chamber 46 Air passage

Claims (10)

Fuel injection means for directly injecting fuel into the cylinder of the engine;
A supply speed control means for variably controlling the supply speed of the air-fuel mixture formed by the fuel injected from the fuel injection means taking in ambient air; and
Combustion starting means for starting combustion of the air-fuel mixture during a fuel injection period;
An operating state detecting means for detecting the operating state of the engine;
Combustion speed detection means for detecting the combustion speed of the air-fuel mixture according to the operating state;
With
It has a first combustion mode in which the supply speed is controlled to be substantially equal to the combustion speed by the supply speed control means, so that the combustion flame of the air-fuel mixture is localized in the cylinder. In-cylinder direct injection internal combustion engine.   The direct injection internal combustion engine according to claim 1, wherein the supply speed control means variably controls the supply speed by adjusting an injection pressure of the fuel injection means.   2. The supply speed control means controls the supply speed so that a position where the combustion flame is fixed is on the fuel spray downstream side from a position where the fuel is substantially in a gas phase. 2. The direct injection type internal combustion engine according to 2.   The said supply rate control means controls a supply rate so that the average equivalence ratio of the spray cross section in the position where the said combustion flame is settled will be about 2 or less, The any one of Claim 1 to 3 characterized by the above-mentioned. The direct injection type internal combustion engine described in 1.   The said supply speed control means makes the said supply speed the highest speed in the range which satisfy | fills the conditions that the said flame is settled in the cylinder of the said engine. The direct injection type internal combustion engine described in 1.   The said supply speed control means controls the said supply speed so that a fuel-injection period becomes substantially constant also when the load of the said engine changes, The any one of Claim 1 to 4 characterized by the above-mentioned. In-cylinder direct injection internal combustion engine.   The in-cylinder direct injection internal combustion engine according to any one of claims 1 to 6, wherein the combustion starting means is means for igniting an air-fuel mixture by spark ignition.   The said fuel injection means is provided with the air supply means which supplies air to a fuel in the inside of the said fuel injection means, or the nozzle hole of the said fuel injection means, The one of Claim 1 to 7 characterized by the above-mentioned. In-cylinder direct injection internal combustion engine. In addition to the first combustion mode, there is a second combustion mode that injects fuel during the intake stroke and ignites an air-fuel mixture diffused in the cylinder near the compression top dead center,
The direct injection internal combustion engine according to any one of claims 1 to 8, wherein the first combustion mode is performed at least in a low rotation / high load operation region. The actual compression ratio variable means capable of variably controlling the actual compression ratio of the engine is provided,
The direct injection internal combustion engine according to claim 9, wherein when the first combustion mode is executed, an actual compression ratio of the engine is set higher than when the second combustion mode is executed.

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