JP2011226446A - Compressed self-ignition internal combustion engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To prevent timing advance of combustion start time and increase in combustion speed during compressed self-ignition combustion due to local temperature increase of air fuel mixture within a combustion chamber.SOLUTION: With respect each cylinder, first and second intake valves 2, 3 are provided. Intake air sucked from the first intake valve 2 to the combustion chamber 1 is oriented to the part in the vicinity of the center of the combustion chamber 1, and intake air sucked from the second intake valve 3 to the combustion chamber 1 is oriented to the part in the vicinity of the peripheral edge of an inner wall of the combustion chamber 1. First and second intake heating means 6, 8 are provided in intake flow passages 7, 9 leading to the first and second intake valves 2, 3, respectively. In accordance with an operating state of an internal combustion engine, heating target temperatures of the first and second intake heating means 6, 8 are controlled to different values. Control is performed, so that by setting a temperature of air fuel mixture in the vicinity of the center of the combustion chamber 1 lower than that in the vicinity of the peripheral edge in an intake stroke, heat transfer is caused in a compression stroke and the temperature of the air fuel mixture within the combustion chamber 1 becomes an approximately uniform temperature enabling self-ignition in an ending period of the compression stroke.

Description

  The present invention relates to a compression self-ignition internal combustion engine, and more particularly to a compression self-ignition internal combustion engine in which a mixture of fuel and air formed in a combustion chamber is self-ignited by a compression action of a piston under at least some operating conditions.

  In a compression self-ignition internal combustion engine, the spark discharge used for ignition in a conventional gasoline combustion internal combustion engine is not required, and the air-fuel mixture formed by mixing fuel and air is compressed by the piston. The auto-ignition temperature is reached, and combustion is started simultaneously and multiple times at a plurality of locations in the combustion chamber space.

  The rise in the mixture temperature due to the compression by the piston is due to the adiabatic compression action. Therefore, in order to obtain a stronger adiabatic compression action and cause the mixture temperature to reach the self-ignition temperature, the compression ratio of the compression self-ignition internal combustion engine is generally set higher than that of the conventional spark ignition internal combustion engine.

  The auto-ignition temperature of gasoline is around 300 ° C., although it varies slightly depending on pressure and mixture concentration. Since it is difficult to raise the temperature of air-fuel mixture containing a large amount of normal temperature only to adiabatic compression, only a part of the hot exhaust gas remains in the combustion chamber or heats the air sucked into the combustion chamber. Therefore, it is necessary to make the temperature of the air-fuel mixture higher than that of a conventional gasoline combustion internal combustion engine.

  In this way, a high-temperature air-fuel mixture can be obtained, but part of it is cooled during the compression period. This is because the internal combustion engine is always cooled. In the internal combustion engine, the wall surface and piston forming the combustion chamber are cooled, and the air-fuel mixture existing near the wall surface and near the piston is deprived of heat.

  The air-fuel mixture whose temperature has been reduced by the heat of the wall surface and the piston is separated from the vicinity of the wall surface and the piston by the flow of the air-fuel mixture in the combustion chamber, and a new uncooled air-fuel mixture flows in the vicinity of the wall surface and the piston. This mixture is also cooled and leaves the wall and piston.

  Thus, while there is always an air-fuel mixture that is cooled in the vicinity of the wall surface and the piston, the cooled air-fuel mixture that has left the wall surface and the vicinity of the piston cools the surrounding uncooled air-fuel mixture, thereby The same temperature. By repeating such a behavior, the temperature of all the air-fuel mixtures in the combustion chamber including the air-fuel mixture in the vicinity of the center of the combustion chamber where heat is not directly taken away by the wall surface or the piston eventually becomes substantially equal.

  Such as when the internal combustion engine is operated at a relatively low speed, such as when the cooling action by the air-fuel mixture cooled by the wall surface or the piston has time to reach the vicinity of the center of the combustion chamber, Since the cooling action sufficiently extends to the air-fuel mixture near the center of the combustion chamber, the air-fuel mixture temperature is easily homogenized throughout the combustion chamber.

  However, when the internal combustion engine is operated at a relatively high rotational speed, the cooling is not performed under the condition that the cooling action by the air-fuel mixture cooled by the wall surface or the piston reaches the vicinity of the center of the combustion chamber. The effect does not sufficiently reach the air-fuel mixture near the center of the combustion chamber, and the air-fuel mixture temperature near the center of the combustion chamber tends to be higher than that in the vicinity.

  Even if the internal combustion engine is operated at a relatively low rotational speed, if the combustion load is high and more air is compressed, the rate at which the temperature of the air-fuel mixture rises due to the adiabatic compression increases, so the cooling effect is combusted. The air-fuel mixture in the vicinity of the center of the chamber does not sufficiently reach the air-fuel mixture, and the temperature of the air-fuel mixture in the vicinity of the center of the combustion chamber tends to be higher than that in the vicinity.

  1 (1) to 1 (3) are diagrams illustrating a process in which the temperature of the air-fuel mixture near the center of the combustion chamber increases. 1 (1) to 1 (3) respectively show the intake stroke end timing, the initial stage of compression, and the timing when compression has further progressed. The upper part (a) of these figures shows the mixture distribution image in the combustion chamber at each time, and the lower part (b) shows the radial temperature distribution image of the combustion chamber at each time. Has been.

  Since fresh air or air-fuel mixture having a uniform temperature in the intake stroke is sucked into the combustion chamber, the temperature of the fresh air or air-fuel mixture in the combustion chamber at the end of the intake stroke becomes uniform as shown in (1).

  As shown in (2), when the compression by the piston is started, the mixture temperature rises uniformly due to the adiabatic compression, but at the same time, the mixture in the vicinity of the combustion chamber wall or piston and the other mixture The heat transfer between is started. However, at this time, as shown in (1), since the temperature of the air-fuel mixture was originally uniform, the heat transfer between the air-fuel mixtures is delayed. That is, first, the heat transfer starts after the temperature of the air-fuel mixture near the combustion chamber wall surface and the piston starts to decrease, and the heat transfer is gradually activated.

  Due to the delay in the activation of heat transfer between the air-fuel mixtures, the air-fuel mixture temperature near the center of the combustion chamber becomes higher than the air-fuel mixture temperature around it.

  As shown in (3), when the compression further proceeds, the temperature difference between the air-fuel mixture near the center of the combustion chamber and the air-fuel mixture near the combustion chamber becomes more prominent, and the air-fuel mixture near the center is relatively early in the compression stroke. Only the auto-ignition temperature is reached and combustion begins. Since the temperature of the air-fuel mixture around this period is still low, a part of the air-fuel mixture does not burn sufficiently, and the combustion start time is advanced, leading to a decrease in combustion efficiency.

  Further, since only the air-fuel mixture near the center of the combustion chamber is heated at an early stage, the combustion speed is locally increased, which causes problems such as generation of noise accompanying combustion vibration.

  For example, Patent Document 1 discloses a technique for diluting the concentration of an air-fuel mixture directed near the center of a combustion chamber under high load operation conditions. It is shown.

JP 2006-348800 A JP 10-339220 A JP 11-182246 A

  In the technology according to Patent Document 1, it is clearly shown that combustion with reduced ignitability due to compression can be obtained by distributing a large amount of air-fuel mixture in the vicinity of the periphery of the combustion chamber where the air-fuel mixture temperature becomes lower under high-load operation conditions. Yes. In other words, it shows that stable compressed self-ignition combustion can be obtained by not distributing the air-fuel mixture in the vicinity of the center of the combustion chamber where the ignitability is likely to be improved by increasing the air-fuel mixture temperature.

  In view of the form of compression self-ignition combustion, it is a fact that this technology has the effect of suppressing the advance of the self-ignition time and the increase in the combustion speed. However, for example, even if only air is supplied near the spatially continuous center in the combustion chamber and air-fuel mixture is supplied near the periphery, the air-fuel mixture in the vicinity of the periphery is affected by the complicated air flow in the combustion chamber. Some will flow near the center. As a result, a very lean air-fuel mixture exists in a part of the space near the center, and even if the air-fuel mixture near the periphery starts to burn, the air-fuel mixture cannot burn sufficiently and burns. There was a problem that the combustion gas was discharged from the combustion chamber together with the remaining gas.

  It is also conceivable that the air-fuel mixture having a concentration in the vicinity of the periphery reaches the vicinity of the center due to the influence of a complicated air flow in the combustion chamber. In this case, there is also a problem that the ignition of the air-fuel mixture in the entire combustion chamber is accelerated by igniting earlier than the air-fuel mixture in the vicinity of the periphery where the air-fuel mixture temperature is low.

  Furthermore, in the technique according to Patent Document 1 in which the premixed gas is directed to a part of the space in the combustion chamber, it is necessary to premix air and fuel upstream of the intake valve, so that fuel is directly injected into the combustion chamber. There is also a problem that it cannot be applied to an in-cylinder direct injection internal combustion engine.

  Although different from the compression self-ignition internal combustion engine, Patent Document 2 discloses a technique capable of suppressing the increase in the temperature of the air-fuel mixture in the vicinity of the center of the combustion chamber.

  In the technique according to Patent Document 2, only the air to be sucked is heated by directing the peripheral edge of the combustion chamber from the auxiliary passage that joins the intake port. In patent document 2, it aims at promotion of atomization of the fuel adhering to the wall surface by this heated air.

  Therefore, Patent Document 2 does not intend to make the temperature of the air-fuel mixture in the combustion chamber substantially uniform and is not intended at all. However, in Patent Document 2, the temperature of the air-fuel mixture becomes substantially uniform. If the timing is an early stage of the compression stroke, it is clear that the air-fuel mixture near the center of the combustion chamber becomes higher than the peripheral portion as the compression further proceeds. Therefore, in order to obtain stable compression self-ignition combustion in which the combustion start timing is not advanced and the increase in the combustion speed is suppressed, ideally, the piston reaches the compression top dead center at the latter stage of the compression stroke. During this period, the temperature of the air-fuel mixture must be uniform and the temperature needs to be the auto-ignition temperature. In order to obtain multipoint simultaneous ignition, which is a feature of compression self-ignition combustion, it is also important that the mixture concentration is uniform.

  In order to satisfy the condition regarding the air-fuel mixture temperature, it is necessary to appropriately control two types of air temperatures at the time of inhalation. However, in the technology according to Patent Document 2, a function for heating the intake air is provided only on one side. For this reason, there is a problem that appropriate control according to the outside air temperature and the operating state of the internal combustion engine, that is, control that makes the entire air-fuel mixture almost the self-ignition temperature in the latter half of the compression stroke cannot be performed.

  As yet another technique, Patent Document 3 discloses a technique of mixing combustion gas only on one side of two intake systems. Since the combustion gas is higher than the outside air temperature, the temperature of the intake air mixed with the combustion gas is higher than the temperature of the intake air not mixed. From this, the temperature distribution of the air or air-fuel mixture in the combustion chamber at the beginning of the compression stroke can make the vicinity of the periphery of the combustion chamber high.

  However, since the intake air on the side where no combustion gas is mixed is at the outside air temperature, there is a problem in that, as in the technique of Patent Document 2, it is impossible to control the entire air-fuel mixture to almost the self-ignition temperature in the latter half of the compression stroke. .

  Furthermore, in the technique according to Patent Document 3, an air-fuel mixture formed by mixing air and fuel mixed with combustion gas and an air-fuel mixture formed by mixing intake air and fuel not mixed with combustion gas are formed. Stratified in the combustion chamber. Even if the entire gas mixture can be brought to almost the auto-ignition temperature in the later stage of the compression stroke, the oxygen concentration in the gas mixture containing the combustion gas is low, so that the self gas is mixed with the gas mixture not containing the gas. There was also a problem that ignition could not be performed and stable compression self-ignition combustion could not be obtained.

  The present invention has been made to solve such a problem, and the local mixture is mixed with the air-fuel mixture in the combustion chamber at the time when the piston reaches the compression top dead center regardless of the outside air temperature or the operating condition of the internal combustion engine. An object of the present invention is to provide a compression self-ignition internal combustion engine that can stabilize the compression self-ignition combustion of the internal combustion engine by obtaining a mixture temperature that does not cause a high temperature and that can self-ignite. And

  The present invention is a compression self-ignition internal combustion engine that self-ignites a mixture of fuel and air formed in a combustion chamber by compression by a piston, and includes first and second intake valves arranged for each cylinder; First and second intake passages for introducing intake air into the combustion chamber via the first and second intake valves, respectively, and the first intake passage and the first intake valve through the first intake valve First directing means for directing the intake air sucked into the combustion chamber to the vicinity of the center of the combustion chamber, and the intake air sucked into the combustion chamber from the second intake passage through the second intake valve Second directing means for directing the vicinity of the periphery of the combustion chamber and the first intake passage, and the intake air flowing through the first intake passage is heated based on a first heating target temperature. Provided in the first intake heating means and the second intake flow path; A second intake air heating means for heating the intake air flowing through the second intake flow path based on a second heating target temperature; and a temperature of the intake air that is directed near the center of the combustion chamber is a peripheral edge of the combustion chamber Based on the current operating state of the internal combustion engine, a temperature difference between the first heating target temperature and the second heating target temperature is set so as to be lower than the temperature of the intake air that is directed in the vicinity, A compression self-ignition internal combustion engine comprising control means for controlling values of the first heating target temperature and the second heating target temperature based on the temperature difference.

  The present invention is a compression self-ignition internal combustion engine that self-ignites a mixture of fuel and air formed in a combustion chamber by compression by a piston, and includes first and second intake valves arranged for each cylinder; First and second intake passages for introducing intake air into the combustion chamber via the first and second intake valves, respectively, and the first intake passage and the first intake valve through the first intake valve First directing means for directing the intake air sucked into the combustion chamber to the vicinity of the center of the combustion chamber, and the intake air sucked into the combustion chamber from the second intake passage through the second intake valve Second directing means for directing the vicinity of the periphery of the combustion chamber and the first intake passage, and the intake air flowing through the first intake passage is heated based on a first heating target temperature. Provided in the first intake heating means and the second intake flow path; A second intake air heating means for heating the intake air flowing through the second intake flow path based on a second heating target temperature; and a temperature of the intake air that is directed near the center of the combustion chamber is a peripheral edge of the combustion chamber Based on the current operating state of the internal combustion engine, a temperature difference between the first heating target temperature and the second heating target temperature is set so as to be lower than the temperature of the intake air that is directed in the vicinity, Since it is a compression self-ignition internal combustion engine comprising control means for controlling the values of the first heating target temperature and the second heating target temperature based on the temperature difference, Regardless of the operating conditions of the engine, at the time when the piston reaches compression top dead center, by obtaining a mixture temperature that does not cause local high temperature in the mixture in the combustion chamber and can self-ignite, Stabilize compression self-ignition combustion of internal combustion engines Door can be.

It is a figure explaining the change of mixture temperature distribution in the compression stroke at the time of suck | inhaling the fresh air of the same temperature via each intake valve in the conventional internal combustion engine. It is a figure explaining the change of mixture temperature distribution in the compression stroke at the time of suck | inhaling the fresh air from which temperature differs through each intake valve in the compression self-ignition internal combustion engine which concerns on Embodiment 1 of this invention. It is a block diagram which shows the structure of the intake system of the compression self-ignition internal combustion engine which concerns on Embodiment 1 of this invention. It is a figure explaining the direction in the combustion chamber of each intake air in the compression self-ignition internal combustion engine which concerns on Embodiment 1 of this invention. It is a figure which shows an example of the means which directs the combustion chamber center to the intake air in the compression self-ignition internal combustion engine which concerns on Embodiment 1 of this invention. It is a figure which shows an example of the means which orient | assigns the periphery of a combustion chamber to the intake air in the compression self-ignition internal combustion engine which concerns on Embodiment 1 of this invention. It is a figure which shows the tendency of each intake air temperature setting with respect to the driving | running state of the internal combustion engine in the compression self-ignition internal combustion engine which concerns on Embodiment 1 of this invention. It is a figure which shows the example of the combustion chamber pressure generation | occurrence | production pattern with respect to the crank angle with or without knocking generation | occurrence | production. It is the figure which represented an example of the combustion chamber pressure signal at the time of knocking generation as spectrum intensity for every frequency. It is a flowchart explaining the flow of control which correct | amends an intake air temperature control target value according to the spectrum intensity of the specific frequency band of the combustion chamber pressure signal in the compression self-ignition internal combustion engine which concerns on Embodiment 2 of this invention.

Embodiment 1 FIG.
FIG. 3 is a diagram showing the configuration of the intake system in the compression self-ignition internal combustion engine according to the first embodiment of the present invention. In FIG. 3, a four-cylinder internal combustion engine is described as an example of a compression self-ignition internal combustion engine, but the number of cylinders is not limited to this. Further, the internal combustion engine according to Embodiment 1 of the present invention is not limited to an internal combustion engine of a type in which an air-fuel mixture in which air and fuel are mixed is sucked into the combustion chamber, but also directly sucks air into the combustion chamber. An internal combustion engine of an injection type is also included. Therefore, in the following description, the air or air-fuel mixture sucked into the combustion chamber is collectively referred to as intake air.

  As shown in FIG. 3, the internal combustion engine according to Embodiment 1 of the present invention includes a combustion chamber 1 provided for each cylinder, a first intake valve 2 and a second intake air arranged for each cylinder. A valve 3, an exhaust valve 4 arranged for each cylinder, a throttle 5 for adjusting an intake amount of intake air sucked into the combustion chamber 1 via the first intake valve 2 and the second intake valve 3, A first intake air heating means 6 that heats one intake air that is divided into two downstream of the throttle 5 and that is sucked into the combustion chamber 1 from the first intake valve 2, and a first end that corresponds to the number of cylinders. A first intake passage 7 that branches off as an intake manifold and is heated by the first intake heating means 6 to a first intake valve 2 provided for each cylinder, and downstream of the throttle 5. The second intake air that is divided into two by the second intake valve 3 and sucked into the combustion chamber 1 from the second intake valve 3 is heated. The intake air heating means 8 and the second intake valve provided for each cylinder are divided into a second intake manifold that branches according to the number of cylinders, and the intake air heated by the second intake air heating means 8 is provided for each cylinder. 2 and a second intake flow path 9 leading to 3. In the configuration of FIG. 3, as the exhaust valve 4, two exhaust valves 4a and 4b are provided for each cylinder.

  In the first embodiment, control means (not shown) for controlling the respective heating target temperatures is provided for the first intake air heating means 6 and the second intake air heating means 8. Yes. The control means responds to the current operating state so that the temperature of the intake air heated by the first intake air heating means 6 and the temperature of the intake air heated by the second intake air heating means 8 have a temperature difference between each other. A temperature difference is set, and values of the first and second heating target temperatures are determined based on the temperature difference. The first intake air heating means 6 and the second intake air heating means 8 individually heat the divided intake air according to the first and second heating target temperatures determined by the control means, respectively. Specifically, the heating target value is determined so that the temperature of the intake air heated by the first intake air heating unit 6 is lower than the temperature of the intake air heated by the second intake air heating unit 8. The Details will be described later.

  In the first embodiment, the first intake passage 7 and the second intake passage 9 are provided with first and second directing means, which will be described later, respectively. The intake air sucked into the combustion chamber 1 from the flow path 7 via the first intake valve 2 is directed near the center of the combustion chamber 1, and is directed from the second intake flow path 9 via the second intake valve 3. The intake air sucked into the combustion chamber 1 is directed to the peripheral edge of the inner wall of the combustion chamber 1 so that the spaces in the combustion chamber 1 to which the respective intake air is directed are different from each other. Details of these directing means will be described later.

  In FIG. 3, configurations such as an intake duct, an air cleaner, an intake port, a fuel injection nozzle, an exhaust port, a crankshaft, and various sensors that are normally provided in a general compression self-ignition internal combustion engine are omitted. However, even in the internal combustion engine of the present invention, those configurations are naturally provided as necessary.

  Next, the operation of the compression self-ignition internal combustion engine according to Embodiment 1 of the present invention will be described. First, the intake operation in the four-cycle internal combustion engine will be described with reference to FIG. In the four-cycle internal combustion engine, four strokes of an intake stroke, a compression stroke, an expansion stroke, and an exhaust stroke are continuously repeated for each combustion chamber 1 of each cylinder. In a four-cylinder internal combustion engine, there are four combustion chambers 1 (hereinafter referred to as combustion chambers 1a to 1d), and an output shaft (not shown) of the crankshaft of the internal combustion engine is provided between the combustion chambers 1a to 1d. These four strokes are set to be repeated with a phase difference of 180 degrees in rotation angle.

  As described above, the first intake valve 2 and the second intake valve 2 for introducing the intake air (air newly sucked into the combustion chamber 1) into the combustion chamber 1 in the intake stroke for each of the combustion chambers 1a to 1d. An intake valve 3 and an exhaust valve 4 for discharging combustion gas from the combustion chamber 1 in the exhaust stroke are provided. In FIG. 3, only the first intake valve 2, the second intake valve 3, and the exhaust valve 4 provided in the combustion chamber 1 d are denoted by symbols, but the other combustion chambers 1 a to 1 c are also illustrated. Thus, each valve 2-4 is provided by the same structure.

  Since each combustion chamber 1 executes the intake stroke with a phase difference of 180 degrees at the rotation angle of the output shaft (not shown) of the crankshaft of the internal combustion engine, any one of the combustion chambers is always operated during the operation of the internal combustion engine. A first intake valve 2 and a second intake valve 3 provided in 1 are opened.

  When the first intake valve 2 and the second intake valve 3 are opened, the piston (not shown in FIG. 3, see FIG. 2) is lowered to increase the volume of the combustion chamber 1 The intake air is sucked into the combustion chamber 1 by the operation.

  The intake air is diverted after passing through the throttle 5 for adjusting the intake amount. One of the divided intake air is heated by the first intake air heating means 6, and further passes through the first intake air passage 7 and the first intake valve 2 to reach the combustion chamber 1. The other divided intake air is heated by the second intake air heating means 8 and further passes through the second intake passage 9 and the second intake valve 3 to reach the same combustion chamber 1.

  The structure of the first intake air heating means 6 and the second intake air heating means 8 is not particularly limited. However, the structure is constituted by an electric heater, or heat capable of exchanging exhaust heat of an internal combustion engine such as combustion exhaust gas with intake air. What consists of an exchanger is desirable.

  The intake air thus divided is heated by the separate intake heating means 6 and 8 and then drawn into the same combustion chamber 1. As described above, the space in the combustion chamber 1 to which each intake air is directed. Are configured in different spaces. FIG. 4 is a diagram for explaining the orientation of each intake air in the combustion chamber 1. In FIG. 4, since each component is the same as FIG. 3, description of a component is abbreviate | omitted here.

  As indicated by a black arrow in FIG. 4, the intake air that has passed through the first intake passage 7 and reached the first intake valve 2 is directed near the center of the combustion chamber 1, while the second intake air The intake air that has passed through the flow path 9 and reached the second intake valve 3 is directed near the periphery of the combustion chamber 1 as indicated by the white arrow. As a result, the intake air that has passed through the first intake passage 7 and reached the first intake valve 2 gathers near the center of the combustion chamber 1 and passes through the second intake passage 9 so as to surround it. The intake air that has reached the second intake valve 3 flows around the periphery of the combustion chamber 1 along the periphery.

  FIG. 5 is a diagram showing an example of a first directing means for directing the vicinity of the center of the combustion chamber 1 to the intake air. In FIG. 5, reference numeral 10 denotes a vertical vortex generating plate as a first directing means, and other components are the same as those in FIG. The base of the vertical vortex generating plate 10 is attached to the inner wall surface on the lower side of the first intake flow path 7, and the tip of the vertical vortex generating plate 10 is rotatable with the base as a fulcrum. The vertical vortex generating plate 10 is moved along the inner wall surface of the flow path 7 or the vertical vortex generating plate 10 is separated from the inner wall surface of the first intake flow path 7 with a predetermined angle. It is possible to do.

  As shown in FIG. 5 (A), in normal intake with the vertical vortex generating plate 10 brought close to the inner wall surface of the first intake flow path 7, the intake is the first intake as shown by the broken line arrow. Since it is sucked into the combustion chamber 1 almost evenly from the entire circumference of the valve 2, there is no special directivity or vortex generation.

  On the other hand, as shown in (B), the front end of the vertical vortex generating plate 10 is lifted from the inner wall surface of the first intake flow path 7 so that the vertical vortex generating plate 10 has a predetermined angle and the first intake flow. In addition to increasing the flow velocity of the intake air, if the vertical vortex generating plate 10 separates the partial flow of the first intake flow path 7 from the inner wall surface of the passage 7 and generates a drift, Intake is sucked into the combustion chamber 1 only from the direction of a part of the intake valve 2, and directivity or strong vortex that directs the vicinity of the center of the combustion chamber 1 to the intake air can be generated. As shown in FIG. 5B, when intake air is drifted to the upper wall surface of the combustion chamber 1 and sucked, most of the intake air collides with the upper wall surface of the combustion chamber 1 and flows toward the center of the combustion chamber 1.

  FIG. 6 is a diagram showing an example of the second directing means for directing the intake air near the periphery of the combustion chamber 1. In FIG. 6, reference numeral 11 denotes a swirl vortex generating plate as the second directing means, and the other components are the same as those in FIG. 3. The base of the swirl vortex generating plate 11 is attached to the inner wall surface on the upper side of the second intake flow path 9, and the tip of the swirl vortex generating plate 11 is rotatable with the base as a fulcrum. The swirl vortex generating plate 11 is moved along the inner wall surface of the path 9, or the swirl vortex generating plate 11 is separated from the inner wall surface of the second intake flow path 9 with a predetermined angle. It is possible.

  As shown in FIG. 6, the tip of the swirl vortex generating plate 11 is lifted from the inner wall surface of the second intake flow path 9, and the swirl vortex generating plate 11 has a predetermined angle and the second intake flow path 9 has a predetermined angle. When the partial flow is generated by blocking the partial flow of the second intake flow path 9 by the swirl vortex generating plate 11 away from the inner wall surface, the flow rate of the intake air is increased and the second intake valve 3 It is possible to obtain directivity in which intake air is sucked into the combustion chamber 1 only from a part of the direction and directs the intake air toward the periphery of the inner wall of the combustion chamber 1. By such means, the intake air sucked in the vicinity of the peripheral edge of the inner wall of the combustion chamber 1 generates a swirl flow along the peripheral edge of the inner wall of the combustion chamber 1.

  When the structure as described above is used, the action of the vertical vortex generating plate 10 and the swirl vortex generating plate 11 generates the intake air directed near the center in the combustion chamber 1 and the intake air directed near the periphery in the combustion chamber 1. In addition, by the action of the first intake air heating means 6 and the second intake air heating means 8, a temperature difference is given to the temperature of the intake air, as shown in (1) of FIG. A temperature difference can be obtained in the air or air-fuel mixture in the combustion chamber 1 at the end of the intake stroke.

  This technique is effective in suppressing the temperature rise of the air or air-fuel mixture in the vicinity of the center of the combustion chamber 1 that is difficult to cool by the wall surface of the combustion chamber 1 and the piston. That is, as shown in FIG. 2, at the end of the intake stroke, the intake temperature sucked near the center of the combustion chamber 1 is set to be lower than the intake temperature sucked near the periphery of the combustion chamber, so that during the compression stroke. It can be controlled so that the temperature of the air-fuel mixture in the combustion chamber 1 is almost uniform at the end of the compression stroke in which the heat has moved and the compression has progressed.

  FIGS. 2 (1) to (3) are diagrams for explaining changes in the mixture temperature in the compression stroke when fresh air having different temperatures is drawn through the intake valves, and the mixture in the vicinity of the center of the combustion chamber. It is a figure explaining the process in which temperature becomes high. FIGS. 2 (1) to 2 (3) show the intake stroke end timing, the initial stage of compression, and the timing when the compression further proceeds. Moreover, the upper part (a) of these figures shows an image of the mixture distribution in the combustion chamber at each time, and the lower part (b) shows an image of the temperature distribution in the radial direction of the combustion chamber at each time. It is shown.

  As shown in FIG. 2 (1), when the intake stroke is completed, air or air-fuel mixture having a temperature higher than that near the center of the combustion chamber is formed around the periphery of the combustion chamber.

  Subsequently, as shown in FIG. 2 (2), the temperature of all the air or air-fuel mixture in the combustion chamber 1 rises as the compression stroke proceeds. At this time, since the air or mixture near the periphery is still hotter than the center of the combustion chamber 1, the heat of the air or mixture near the periphery moves to the air or mixture near the center.

  As shown in FIG. 2 (3), while the heat is transferred, when the compression further proceeds, the entire air-fuel mixture becomes a substantially homogeneous temperature.

  Thereby, in the internal combustion engine according to the present embodiment, the temperature of the air-fuel mixture in the combustion chamber 1 becomes uniform at the later stage of the compression stroke, ideally at the time when the piston reaches the compression top dead center. The temperature can be controlled to be the auto-ignition temperature. Further, in order to obtain multipoint simultaneous ignition, which is a feature of compression self-ignition combustion, it is important that the mixture concentration in the combustion chamber 1 is uniform. In the internal combustion engine according to the present embodiment, however, Since the concentration of the air-fuel mixture sucked from the first intake valve 2 and the second air-intake valve 3 is uniform, the air-fuel mixture concentration in the combustion chamber 1 does not vary and the air-fuel mixture concentration in the combustion chamber 1 does not vary. Is naturally homogeneous. As described above, in the present embodiment, it is possible to realize stable compression self-ignition combustion in which the combustion start time is not advanced and the increase in the combustion speed is suppressed.

  FIG. 7 shows each intake air temperature setting pattern with respect to the operating state of the internal combustion engine (shaft rotation speed, intake air amount, etc.). In the graph shown in FIG. 7, the horizontal axis represents the rotational speed of the internal combustion engine and the intake air amount, and the vertical axis represents the temperature of each intake air.

  A characteristic line indicated by a solid line in FIG. 7 indicates the temperature of the intake air that is directed near the periphery of the combustion chamber 1, and indicates that the intake air temperature is decreased as the shaft rotational speed and the intake air amount of the internal combustion engine increase. I mean.

  This is because the time for heat transfer from the air or mixture in the combustion chamber 1 to the wall surface or piston of the combustion chamber 1 is shortened as the shaft rotational speed of the internal combustion engine increases, so that the adiabatic compression heat is converted into the combustion chamber 1. This is because the temperature rise of the air and the air-fuel mixture due to adiabatic compression increases as the intake air amount increases.

  In addition, the characteristic line indicated by a broken line in FIG. 7 indicates the temperature of the intake air directed near the center of the combustion chamber 1, and the combustion chamber indicated by the solid line as the shaft rotational speed and intake air amount of the internal combustion engine increase. This means that the temperature difference ΔT with the temperature of the intake air directed to the vicinity of the peripheral edge of 1 is enlarged.

  This is because the time during which the heat between the air and the air-fuel mixture in the combustion chamber 1 moves with the increase in the shaft speed of the internal combustion engine is shortened, so that the heat of the air and air-fuel mixture near the center of the combustion chamber 1 is reduced. Because it is difficult for the air and air-fuel mixture near the periphery of the combustion chamber 1 to be deprived, and as the intake air amount increases, the temperature rise itself due to adiabatic compression of the air and air-fuel mixture near the center of the combustion chamber 1 increases. It is.

  The specific temperature setting of the temperature difference ΔT corresponding to the increase or decrease of the shaft rotational speed and the intake air amount that is the current operating state of the internal combustion engine is the difference in the characteristics of the internal combustion engine, particularly the structure of the combustion chamber 1, for example, It differs depending on the diameter of the combustion chamber 1, the piston shape, and the difference in the ability to cool the air-fuel mixture. Therefore, in the present embodiment, for each internal combustion engine having a different structure of the combustion chamber 1, a predetermined function relating to the shaft rotational speed and intake air amount for deriving the temperature difference ΔT for each shaft rotational speed and intake air amount, A set value data table is created in advance.

  Further, a second intake flow path 9 from the first intake heating means 6 to the first intake valve 2 and a second intake flow path 9 from the second intake heating means 8 to the second intake valve 3 is provided. Each is provided with an intake air temperature detection means (not shown) composed of a thermocouple, a thermistor, etc., and each intake air temperature is obtained from the intake air temperature detection means, and the functions and settings described above are obtained accordingly. Based on the temperature difference ΔT obtained from the value data table, the heating amount control by each of the first intake air heating means 6 and the second intake air heating means 8 is performed.

  Specifically, first, the intake air temperature difference ΔT according to the current operating state (shaft rotation speed or intake air amount) is calculated using the above-described function or set value data table preset for each characteristic of the internal combustion engine. Find and set. Next, for the heating target temperature T2 of the intake air sucked through the second intake valve 3 so as to turn around the periphery of the combustion chamber 1, the shaft rotational speed is similarly determined for each characteristic of the internal combustion engine. Further, a predetermined function and set value data table relating to the intake air amount are set in advance, and the heating target temperature T2 is obtained from the function and set value data table and set. Next, by subtracting the temperature difference ΔT from the heating target temperature T2, the heating target temperature T1 of the intake air sucked through the first intake valve 2 that is directed near the center of the combustion chamber 1 is calculated (T1 = T2). -ΔT). The heating target temperature T1 thus obtained is set as the heating target temperature of the intake air by the first intake air heating means 6, and the heating target temperature T2 is set as the heating target temperature of the intake air by the second intake air heating means 8. Note that this control flow is repeatedly executed at a predetermined cycle, thereby making it possible to cope with the current operating state of the internal combustion engine that changes from moment to moment. This makes it possible to control the heating amount by each of the first intake air heating means 6 and the second intake air heating means 8 based on the temperature difference ΔT set according to the current operating state, and directs the vicinity of the center of the combustion chamber. The temperature of the intake air to be controlled can be controlled to be lower by the temperature difference ΔT set according to the current operating state than the temperature of the intake air directed toward the periphery of the combustion chamber.

  When both the intake air is heated by the first intake air heating means 6 and the second intake air heating means 8, low-temperature air is not directly sucked into the combustion chamber 1, so that the air and the air-fuel mixture at the end of the compression stroke are always self. The temperature can be ignited, and the first intake air heating means 6 and the second intake air heating means 8 can always appropriately control the temperature difference of the intake air based on the value of the temperature difference ΔT. The temperature of the air-fuel mixture in the combustion chamber 1 at the end time of can be made uniform. As a result, while preventing misfire due to the low temperature of the air-fuel mixture, the self-ignition caused by the local high temperature in the air-fuel mixture in the combustion chamber 1 can be accelerated, and the compression self-ignition combustion accompanying the increase in the combustion speed can be prevented. Instability is prevented.

  As described above, according to the compression self-ignition internal combustion engine according to the first embodiment of the present invention, using the first and second intake air heating means 6 and 8 separately provided in the two intake systems, In accordance with the operating state of the internal combustion engine, at the end of the intake stroke, the intake air temperature set so as to be directed to the vicinity of the center of the combustion chamber 1 is set in advance so as to swirl toward the periphery of the combustion chamber 1. The temperature is controlled to be lower than the temperature, and heat is transferred during the compression stroke, so the end of the compression stroke (the timing at which the piston reaches the compression top dead center regardless of the operating state of the internal combustion engine). ), The entire air-fuel mixture in the combustion chamber 1 can be controlled to a substantially uniform temperature. Also, by controlling the heating of both intake air, the air-fuel mixture in the combustion chamber 1 can be brought to a temperature at which self-ignition is possible at a time when the piston reaches compression top dead center, regardless of the outside air temperature or the like. Early self-ignition caused by local high temperature in the air-fuel mixture in the combustion chamber 1, destabilization of compression self-ignition combustion accompanying an increase in combustion speed, and the air-fuel mixture temperature in the combustion chamber 1 is self The unburned mixture is prevented from being discharged due to misfires caused by the inability to reach an ignitable temperature. Thus, according to the compression self-ignition internal combustion engine according to the first embodiment, the air-fuel mixture in the combustion chamber 1 is approximately the time when the piston reaches the compression top dead center regardless of the outside air temperature or the operating conditions of the internal combustion engine. Therefore, it is possible to obtain a mixture temperature that can be self-ignited without causing local high temperature, and to stabilize the compression self-ignition combustion of the internal combustion engine. Thus, it is possible to prevent the unburned mixture from being discharged due to unstable combustion.

Embodiment 2. FIG.
In the first embodiment described above, the temperature of the intake air directed to the vicinity of the center of the combustion chamber 1 and the vicinity of the periphery of the combustion chamber 1 set in advance for each shaft speed and intake air amount in accordance with the characteristics of the internal combustion engine. The operation has been described in which the amount of heating of each intake air is controlled based on the temperature difference ΔT with the temperature of the intake air to prevent the early ignition and the instability of the compression self-ignition combustion. In the second embodiment, the temperature difference ΔT is corrected based on the result of detecting the actual combustion state, thereby more reliably preventing the early self-ignition and the instability of the compression self-ignition combustion. The control operation will be described.

  The configuration of the compression self-ignition internal combustion engine according to the second embodiment of the present invention is basically the same as the configuration of the first embodiment shown in FIGS. The description will be omitted.

  FIG. 8 is a diagram showing an example of a combustion chamber pressure generation pattern with respect to a crank angle with or without knocking. The thick solid line in FIG. 8 shows an example of the pressure change in the combustion chamber 1 with respect to the crank angle when an abnormality occurs in compression self-ignition due to the local high temperature of the air-fuel mixture in the combustion chamber 1. The thin solid line indicates an example of the pressure change when there is no abnormality due to local high temperature of the air-fuel mixture. The abnormality in this case is a state in which a high-frequency pressure change component is added to a normal pressure change as in so-called knocking.

  When a high-frequency change such as knocking is added to the pressure change in the combustion chamber, the drive torque transmitted to the output shaft (not shown) of the crankshaft of the internal combustion engine vibrates in small increments. The running speed changes in small increments and causes noise, which impairs the comfort of the vehicle.

  In a compression self-ignition internal combustion engine, the main factor that a high-frequency change such as knocking is added to such a pressure change is the local high temperature of the air-fuel mixture, and the local high temperature is caused by the wall surface of the combustion chamber 1 and the piston. Since it occurs in the vicinity of the center of the combustion chamber 1 where the cooling effect is difficult to achieve, control for expanding the temperature difference ΔT shown in FIG. 7 is effective in preventing this.

  FIG. 9 is a diagram showing the spectral intensity of each signal frequency obtained by Fourier transforming the pressure change data with respect to the crank angle under the condition that a high frequency change is applied to the pressure change in the combustion chamber 1 like knocking. .

  The increase in the spectrum on the low frequency side is a change caused by a basic pressure change due to compression self-ignition combustion that is obtained even when a high frequency change is not applied like knocking.

  On the other hand, the increase in the spectrum generated on the high frequency side is due to the generation of a high frequency component such as knocking applied to the pressure change. Therefore, by obtaining the spectrum (its intensity) in this high frequency band, it is possible to detect the presence and intensity of the high frequency component as in the knocking applied to the pressure change.

  Such spectral intensity P in the high frequency band is obtained by arithmetically calculating the output of the element capable of converting pressure change information in the combustion chamber 1 and some physical quantity change linked to the pressure change into a voltage signal. However, the spectrum intensity of the specific frequency band can be obtained by a method such as using a filter that passes only a specific high frequency band in the output signal of the element. Since the high frequency component is generated by resonance of energy that moves at high speed in the combustion chamber, the shape of the combustion chamber 1 is dominant in the frequency. Therefore, it is not necessary to change the frequency band for detecting the spectrum intensity according to the frequency of the internal combustion engine operation such as the shaft rotational speed.

  When the spectrum intensity P in this specific frequency band is larger than, for example, a predetermined upper limit value, the temperature setting value of the intake air sucked through the first intake valve 2 that is directed near the center of the combustion chamber 1 is set from the current state. Further, by correcting the value to a low value and executing the intake air heating control, it is possible to suppress the occurrence of shaft torque vibration and noise of the internal combustion engine.

  An example of a specific processing flow is shown in FIG. In this example, control is performed so that the spectrum intensity in the high frequency band related to the pressure information in the combustion chamber becomes a preset value.

  From the start of control, first, in step S100, the spectrum intensity control target value Ps for the high frequency band is set in a memory (not shown). The spectrum intensity control target value Ps is an upper limit value of the spectrum intensity allowed by the internal combustion engine, and may be any value that is greater than 0 and less than or equal to the upper limit value.

  Next, in step S101, the second intake air is swung around the temperature difference ΔT0 of the intake air corresponding to the current operating state of the internal combustion engine (shaft rotation speed, intake air amount, etc.) and the periphery of the combustion chamber 1. The heating target temperature T2 of the intake air sucked through the valve 3 is set by calculation from a predetermined function set in advance, or is set with reference to a preset table.

  Next, in step S102, first, the current spectrum intensity P is detected based on the pressure information in the combustion chamber 1. As a detection method, any one of the above-described methods may be used. For example, an output of an element that can obtain an electrical output corresponding to a pressure change in the combustion chamber 1 is obtained by arithmetic calculation, or The output of the element is obtained by passing through a filter that passes only a specific high frequency band. Next, the control coefficient based on the deviation between the detected current spectral intensity P and the spectral intensity control target value Ps set in the memory in step S100 is executed to calculate the correction coefficient R (R = f (P− Ps), 0 ≦ R). The control calculation includes proportional control calculation (P control) for deviation, PI control in which integral control calculation (I control) proportional to the integral value of deviation is added to proportional calculation, and proportional calculation is proportional to changes in spectrum intensity. PD control to which differential control calculation (D control) is added, PID control that combines these, and the like can be used.

  In step S103, the intake air temperature difference ΔT is calculated by multiplying the intake air temperature difference ΔT0 according to the current operating state of the internal combustion engine set in step S101 by the correction coefficient R obtained in step S102 (ΔT = ΔT0 × R).

  Next, in step S104, the temperature difference ΔT is determined from the heating target temperature T2 of the intake air sucked through the second intake valve 3 so as to turn near the periphery of the combustion chamber 1 set in step S101. By subtracting, the heating target temperature T1 of the intake air sucked through the first intake valve 2 directed near the center of the combustion chamber 1 is calculated (T1 = T2−ΔT).

  Therefore, it is desirable that the correction coefficient R obtained as a result of the control calculation in step S102 is a value of 0 or more. This is because when R is a negative value, the temperature of the intake air that is directed near the center of the combustion chamber 1 becomes higher than the temperature of the air that is sucked so as to swirl around the periphery of the combustion chamber 1, and the compression stroke. This is because the local temperature increasing tendency of the air-fuel mixture temperature near the center of the combustion chamber 1 increases with the progress of.

  Next, in step S105, the heating target temperature T1 of the intake air that is directed near the center of the combustion chamber 1 calculated in step S104 is set as the heating target temperature of the intake air by the first intake air heating means 6, and in step S101. The target heating temperature T2 of the intake air that is directed along the peripheral edge of the combustion chamber 1 is set as the target heating temperature of the intake air by the second intake air heating means 8.

  This control flow is repeatedly executed by returning from step S105 to step S101. Therefore, the heating target temperature of the intake air by each of the first intake air heating means 6 and the second intake air heating means 8 changes every moment according to the change in the operating state of the internal combustion engine. For this reason, in the present embodiment, the first intake passage 7 from the first intake heating means 6 to the first intake valve 2 and the second intake heating means 8 from the second intake heating means 8 to the second intake valve 3 are reached. Each intake air temperature is obtained from an intake air temperature detecting means (not shown) such as a thermocouple or thermistor provided in each of the second intake air passages 9, and the temperature difference from the heating target temperature T1 or the heating target temperature T2 or the like. By continuously executing the heating control by the first intake air heating means 6 or the second intake air heating means 8 based on the above, the temperature difference of the intake air is always kept appropriate in the internal combustion engine during the combustion operation. In the compression self-ignition internal combustion engine, the temperature of the air-fuel mixture in the combustion chamber does not cause local high temperature and the self-ignition temperature is improved. Early ignition and increased combustion speed Compression destabilization of the self-ignition combustion can be prevented with.

  As described above, according to the second embodiment, the same effect as in the first embodiment described above can be obtained, and furthermore, in the second embodiment, an electrical response corresponding to the pressure change in the combustion chamber can be obtained. The spectrum intensity P in the high frequency band specified from the output of the element from which the output is obtained is detected, and the heating target temperature T1 of the air or the air-fuel mixture that is directed near the center of the combustion chamber 1 according to the spectrum intensity P and the combustion chamber 1 Because the temperature difference ΔT with the target heating temperature T2 of the air or air-fuel mixture that runs along the periphery of the engine is corrected, the self-ignition and the combustion speed increase more reliably regardless of the operating state of the internal combustion engine. The effect of preventing the instability of compression self-ignition combustion and the accompanying discharge of unburned mixture is obtained.

  The operation according to the first and second embodiments of the present invention is described for the case where air having different temperatures is sucked into the combustion chamber, and relates to a compression self-ignition internal combustion engine of a direct injection type. . However, the fluid sucked into the combustion chamber through one intake valve is directed to the center of the combustion chamber, and the temperature is controlled to be lower than the fluid sucked by directing the periphery of the combustion chamber through the other intake valve, In addition, if the fluid to be sucked is an air-fuel mixture having substantially the same concentration, the same effect can be obtained in a premixed compression self-ignition internal combustion engine.

  DESCRIPTION OF SYMBOLS 1 Combustion chamber, 2 1st intake valve, 3nd intake valve, 4 Exhaust valve, 5 Throttle, 6 1st intake heating means, 7 1st intake flow path, 8 2nd intake heating means, 9 Second intake flow path, 10 longitudinal vortex generator plate, 11 swirl vortex generator plate.

Claims (2)

A compression self-ignition internal combustion engine that self-ignites a mixture of fuel and air formed in a combustion chamber by compression by a piston,
First and second intake valves arranged for each cylinder;
First and second intake passages for introducing intake air into the combustion chamber via the first and second intake valves, respectively;
First directing means for directing the intake air sucked into the combustion chamber from the first intake passage through the first intake valve in the vicinity of the center of the combustion chamber;
Second directing means for directing the intake air sucked into the combustion chamber from the second intake flow path via the second intake valve in the vicinity of the periphery of the combustion chamber;
A first intake air heating means provided in the first intake flow path for heating the intake air flowing through the first intake flow path based on a first heating target temperature;
A second intake air heating means provided in the second intake air flow path for heating the intake air flowing through the second intake flow path based on a second heating target temperature;
Based on the current operating state of the internal combustion engine, the first intake air that is directed near the center of the combustion chamber is lower in temperature than the intake air that is directed near the periphery of the combustion chamber. Control means for setting a temperature difference between the heating target temperature and the second heating target temperature and controlling the values of the first heating target temperature and the second heating target temperature based on the temperature difference. A compression self-ignition internal combustion engine characterized by the above. The control means further includes
A spectrum intensity in a specific high frequency band is detected from an output of an element that can obtain an electrical output corresponding to a pressure change in the combustion chamber, and the first heating target temperature and the second heating are detected according to the spectrum intensity. The temperature difference from the target temperature is corrected, and the values of the first heating target temperature and the second heating target temperature are controlled based on the corrected temperature difference. Compression self-ignition internal combustion engine.

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