JP2012207627A - Gasoline engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a gasoline engine capable of achieving compression self ignition combustion and promoting activity of a catalyser by raising a geometrical compression ratio while avoiding abnormal combustion.SOLUTION: Negative pressure is generated in an exhaust port 10 connected to the other neighboring independent exhaust passage 52 by ejector effect as a shape in which the downstream side thereof makes the downstream end of the independent exhaust passage 52 connected to the exhaust port 10 a smaller flow passage area. In a low load and low speed region, a self ignition combustion mode in which mixture burns by self ignition is executed, and in a high load and low speed region, a valve opening period of an intake valve 11 and a valve opening period of an exhaust valve 12 are overlapped in an overlapped period. The overlapped period of one cylinder 2 among cylinders in continuous exhaust order is overlapped on an exhaust valve 12 opening period of the other cylinder 2.

Description

  The present invention includes a plurality of cylinders each having a combustion chamber, an intake port and an exhaust port, and provided with an intake valve capable of opening and closing the intake port and an exhaust valve capable of opening and closing the exhaust port. The present invention relates to a gasoline engine having an injector for injecting fuel made of gasoline into the combustion chamber.

  Conventionally, in the gasoline engine field, a combustion mode (spark ignition combustion) in which an air-fuel mixture is forcibly ignited by spark ignition of a spark plug has been common, but in recent years, instead of such spark ignition combustion, so-called Research is underway to apply compression auto-ignition combustion to gasoline engines. The compression self-ignition combustion is to compress the air-fuel mixture generated in the combustion chamber with a piston and to ignite the air-fuel mixture in a high temperature / high pressure environment regardless of spark ignition. Compressed self-ignition combustion is combustion in which self-ignition occurs at various points in the combustion chamber at the same time, and it is said that high thermal efficiency can be obtained compared to combustion by spark ignition.

  As a specific example of the gasoline engine to which the compression self-ignition combustion is applied, for example, one disclosed in Patent Document 1 below is known. In Patent Document 1, the geometric compression ratio of the cylinder is set to 14 or more. It is disclosed that the air-fuel mixture is raised to a temperature at which self-ignition is possible.

  Here, when the geometric compression ratio is set to a high value of 14 or more, the temperature of the air-fuel mixture in the combustion chamber rises excessively, and as a result, abnormal combustion such as knocking occurs in the low speed and high load region. There's a problem.

  On the other hand, by using a so-called 4-2-1 type exhaust system as disclosed in Patent Document 2 below, scavenging of the combustion chamber is promoted to reduce the residual gas in the combustion chamber. It has been studied to suppress an excessive temperature rise of the air-fuel mixture. In this exhaust system, in a four-cylinder engine, after four exhaust passages connected to the exhaust ports of each cylinder are gathered two by two, these two exhaust passages are gathered into one, and the exhaust gas is By gradually gathering, the influence of back pressure and exhaust pulsation is suppressed, and scavenging is promoted.

JP 2007-154859 A JP 2002-276356 A

  In the 4-2-1 type exhaust system, the occurrence of abnormal combustion can be suppressed by improving the scavenging performance. On the other hand, as a result of the longer exhaust passage, the exhaust temperature flowing into the catalyst provided in the exhaust passage decreases. As a result, there is a problem that it takes time to activate the catalyst or it becomes difficult to maintain the active state of the catalyst.

  In view of such circumstances, the present invention provides a gasoline engine that can achieve self-ignition combustion by increasing the geometric compression ratio while avoiding abnormal combustion, and can promote the activity of the catalyst. With the goal.

  In order to solve the above problems, the present invention includes a combustion chamber, an intake port and an exhaust port, and an intake valve capable of opening and closing the intake port and an exhaust valve capable of opening and closing the exhaust port. A gasoline engine comprising a plurality of cylinders and an injector for injecting fuel, at least part of which is made of gasoline, into the combustion chamber, and controls the operation of the intake and exhaust valves and the combustion mode in the combustion chamber Control means, independent exhaust passages connected to the exhaust ports of one cylinder or a plurality of cylinders whose exhaust order is not mutually continuous, and the independent exhaust passages so that the gas passing through the independent exhaust passages gathers. A stoichiometric air-fuel ratio atmosphere arranged at a downstream side of the collecting portion connected to the downstream end and the collecting portion so that the gas that has passed through the collecting portion flows in. And a catalyst having a three-way catalyst function, the geometric compression ratio of the cylinder is set to 14 or more, and an independent exhaust passage connected to a cylinder in which the exhaust order is continuous among the independent exhaust passages is Connected to the collecting portion at positions adjacent to each other, and the downstream end portion of each independent exhaust passage is exhausted from each cylinder to the collecting portion through each exhaust port and each independent exhaust passage. Accordingly, the downstream side has a shape in which the flow path area is smaller than the upstream side so that negative pressure is generated in the adjacent independent exhaust passage and the exhaust port connected to the independent exhaust passage due to the ejector effect. The control means executes a self-ignition combustion mode in which the air-fuel mixture in the combustion chamber burns by self-ignition at least in a low load and low speed region of the engine, and at least high engine negative In addition, in the low speed range, the opening period of the intake valve and the opening period of the exhaust valve of each cylinder overlap with each other by a predetermined overlap period, and the overlap period of one cylinder is between cylinders in which the exhaust order is continuous. A gasoline engine is provided in which the intake valve and the exhaust valve of each cylinder are opened so as to overlap with the time when the exhaust valve of the other cylinder is opened.

  According to the present invention, high scavenging performance can be realized in a low-speed and high-load region without increasing the length of the exhaust passage from the exhaust port to the catalyst. Therefore, while avoiding abnormal combustion, it is possible to achieve self-ignition combustion of the air-fuel mixture by setting the geometric compression ratio of the cylinder to 14 or higher, thereby realizing high thermal efficiency and increasing the temperature of the exhaust gas flowing into the catalyst. Thus, the activity of the catalyst can be promoted.

  Specifically, in the present invention, the downstream end portion of each independent exhaust passage has a shape in which the flow area is smaller on the downstream side than on the upstream side, and a high speed is provided from the predetermined independent exhaust passage to the collecting portion. Exhaust gas flows in, and thereby an ejector effect is obtained, and the other independent exhaust passages are configured to have a negative pressure. Since the valve is open, in this low speed and high load region, a negative pressure is generated in the exhaust port of the cylinder during the overlap period by the ejector effect, and scavenging during the overlap period is promoted by this negative pressure. Can do. Therefore, even if the geometric compression ratio is increased, abnormal combustion can be more reliably avoided in the low-speed and high-load region, and proper compression self-ignition combustion associated with the high geometric compression ratio is achieved. Can be realized.

  In addition, high scavenging performance is realized by using the ejector effect, and there is no need to increase the length of the exhaust passage (reducing the length of the exhaust passage compared to the 4-2-1 type exhaust system). The amount of heat released from the exhaust gas until reaching the catalyst can be kept small, the temperature of the exhaust gas flowing into the catalyst can be increased, and the activity of the catalyst can be promoted. In particular, even during compression auto-ignition combustion, where the combustion temperature is low and the exhaust temperature is low, the catalyst is difficult to activate, the activity of the catalyst can be promoted, and the catalyst can be activated while obtaining high thermal efficiency. . In addition, as described above, the exhaust flows from the independent exhaust passage toward the collecting portion at a high speed, so that the exhaust is suppressed from expanding around the other independent exhaust passage, and the temperature of the exhaust decreases due to the expansion. Is also suppressed.

  In the present invention, it is preferable that the collecting portion has a shape in which the flow path area is smaller on the downstream side than on the upstream side (Claim 2).

  If it does in this way, the flow velocity of the exhaust gas which passes through the above-mentioned gathering part can be raised, and the scavenging performance can be further enhanced by increasing the amount of negative pressure generated in the exhaust port.

  In the present invention, it is preferable that a connecting portion having a shape in which the flow path area is larger on the downstream side than on the upstream side is provided between the collecting portion and the catalyst.

  As described above, the exhaust gas passes through the collecting portion at a high speed, and the temperature of the exhaust gas passing through the collecting portion decreases as the flow velocity increases. Therefore, if a connecting portion having such a structure is provided, the flow rate and temperature of the exhaust gas can be restored at this connecting portion (the flow rate decreases and the temperature increases), and the temperature of the exhaust gas flowing into the catalyst is increased. Can be maintained. Further, as the temperature of the exhaust gas decreases, the amount of heat released to the outside at the gathering portion can be kept small. As described above, according to this configuration, the temperature of the exhaust gas can be returned to a high level in the connecting portion while suppressing the amount of heat released from the collecting portion, and compared with the case where the flow area of the connecting portion is constant, The temperature of the inflowing exhaust gas can be increased.

  In the present invention, the control means includes the intake valve of each cylinder not only in a high load and low speed region of the engine but also in at least a partial region of the low load and low speed region of the engine that executes the auto-ignition combustion mode. Between the cylinders in which the valve opening period and the exhaust valve opening period overlap with each other by a predetermined overlap period and the exhaust sequence continues, the exhaust valve of the other cylinder opens during the overlap period of one cylinder. It is preferable to open the intake valve and the exhaust valve of each cylinder so as to overlap with each other at the same time.

  In this way, even in the auto-ignition combustion mode region where the thermal efficiency is high and the exhaust temperature is low, the exhaust of the cylinder during the overlap period is sucked downstream due to the ejector effect, and this exhaust goes to other independent exhaust passages. Since it is avoided more reliably, the expansion of the exhaust gas due to the wraparound can be avoided, and the temperature drop of the exhaust gas can be suppressed. This further promotes the activity of the catalyst.

  In the present invention, the control means can control G / F, which is a ratio of the weight G of the total gas in the combustion chamber to the fuel weight F injected into the combustion chamber by the injector, and the self-ignition combustion The ratio G / F of the total gas weight G in the combustion chamber to the fuel weight F is controlled to 30 or more in the low load and low speed range of the engine executing the mode, and the higher the load side, the higher the load G in the combustion chamber to the fuel weight F. It is preferable to reduce the ratio G / F of the total gas weight G (Claim 5).

  Thus, if compression auto-ignition combustion is executed in a lean state where the air-fuel ratio of the combustion chamber is very high, generation of NOx can be more reliably suppressed, and higher thermal efficiency can be obtained by lowering the combustion temperature. Can do. Here, the temperature of the exhaust gas decreases as the combustion temperature decreases. However, in the present invention, the exhaust gas temperature flowing into the catalyst is increased as described above, so that the activity of the catalyst is achieved while realizing proper compression self-ignition combustion. Can be promoted. Further, in this configuration, the ratio G / F of the total gas weight F in the combustion chamber to the fuel weight F is reduced toward the higher load side, and the appropriate G / F is set according to the increase in the fuel amount. The engine torque can be secured while realizing the compression self-ignition combustion.

  Further, in the present invention, the control means can control the operation of the injector, and in the low load and low speed range of the engine that executes the auto-ignition combustion mode, the injector is in the first half of the intake stroke or the compression stroke. Preferably, fuel is injected into the combustion chamber.

  In this way, since the fuel injected into the combustion chamber is sufficiently mixed with air until the self-ignition occurs, the mixture concentration in the combustion chamber is homogenized, so that the combustion temperature can be lowered more reliably. Thereby, higher thermal efficiency can be obtained. Here, although the exhaust gas temperature decreases as the combustion temperature decreases, in the present invention, as described above, the exhaust gas temperature flowing into the catalyst is increased, so that the catalyst activity is promoted while realizing high thermal efficiency. Can do.

  In the present invention, an ignition plug for supplying ignition energy to the air-fuel mixture in the combustion chamber is provided, and the control means is capable of controlling the operation of the ignition plug and performs the self-ignition combustion mode. A fuel injection for injecting a large amount of fuel from the injector at an injection pressure of 30 MPa or more to make the average air-fuel ratio in the combustion chamber richer than in the self-ignition combustion mode in a region of a higher load than the ignition plug; Is performed within a period from the late stage of the compression stroke to the early stage of the expansion stroke, whereby the air-fuel mixture based on the combustion injection is rapidly burned by flame propagation after the compression top dead center has passed a predetermined period or more. It is preferable to execute the rapid retarded SI mode (claim 7).

  In this way, fuel injection at a high pressure of 30 MPa or more is performed on the retarded side, such as within the period from the late stage of the compression stroke to the early stage of the expansion stroke, and the compression top dead center where the combustion chamber is at the highest temperature and pressure Until a certain amount is passed, a large amount of fuel can be sufficiently vaporized and atomized to form a relatively homogeneous (or weakly stratified) mixture while maintaining a large turbulent energy due to a high injection pressure. In this state, by starting the flame propagation combustion based on the spark ignition of the spark plug, the air-fuel mixture based on the high-pressure injection can be burned out in a short period after the compression top dead center. For this reason, unlike conventional spark ignition combustion, in which fuel is injected at an early timing such as the intake stroke, abnormal combustion such as pre-ignition and knocking can be reliably avoided, and flame propagation combustion with a short combustion period and excellent thermal efficiency can be achieved. Can be realized. Thus, according to this configuration, in addition to the suppression of abnormal combustion accompanying the improvement of the scavenging performance by the ejector effect, it is possible to suppress abnormal combustion depending on the combustion mode, and more reliably generate abnormal combustion. This can be avoided and high thermal efficiency can be obtained in a wider operating range. Further, in this combustion mode, the combustion temperature does not rise excessively, and the combustion does not start with insufficient vaporization and atomization of the fuel. Well maintained.

  As described above, according to the present invention, it is possible to achieve self-ignition combustion by increasing the geometric compression ratio while avoiding abnormal combustion, and promote the activity of the catalyst.

It is a figure showing the whole gasoline engine composition concerning an embodiment of the present invention. It is a figure which shows the structure of the exhaust system of the engine shown in FIG. It is a schematic sectional drawing which shows the shape of the piston crown surface of the engine shown in FIG. FIG. 3 is a schematic side view of the exhaust system shown in FIG. 2. It is the VV sectional view taken on the line of FIG. It is a figure for demonstrating the valve timing of an intake valve and an exhaust valve in a rapid retarded SI combustion mode. It is a figure for demonstrating the effect of the exhaust system shown in FIG. It is a block diagram which shows the control system of the engine shown in FIG. It is a figure which shows an example of the control map for selecting the control mode according to the driving | running state of an engine. FIG. 10 is a time chart for illustrating the control contents of a lean HCCI combustion mode executed in the first operation region A1 of FIG. FIG. 10 is a time chart for illustrating the control content of a rapid retarded SI combustion mode executed in the second operation region A2 of FIG. It is a figure for demonstrating a mode that the fuel injected in the late stage of the compression stroke diffuses in a cavity in rapid retarded SI combustion mode. It is a figure for demonstrating the valve timing of an intake valve and an exhaust valve.

(1) Overall Configuration of Engine FIG. 1 is a diagram showing an overall configuration of a gasoline engine according to an embodiment of the present invention. FIG. 2 is a view showing a part (mainly an exhaust system) of this engine. The engine shown in FIG. 1 is a reciprocating piston type multi-cylinder gasoline engine mounted on a vehicle as a power source for driving driving. The engine body 1 of this engine can reciprocately slide in each cylinder 2 with a cylinder block 3 having four cylinders 2 (see FIG. 2) arranged in a predetermined direction, a cylinder head 4 provided on the upper surface of the cylinder block 3 And the piston 5 inserted into the. Specifically, a first cylinder 2a, a second cylinder 2b, a third cylinder 2c, and a fourth cylinder 2d are formed in order from the right in FIG. In addition, the fuel supplied to the engine body 1 may be anything that contains gasoline as a main component, and the contents may be all gasoline, or gasoline containing ethanol (ethyl alcohol) or the like. But you can.

  The engine body 1 is a four-cycle engine, and is configured such that the intake stroke, the compression stroke, the expansion stroke, and the exhaust stroke are shifted by 180 ° C. in each of the cylinders 2a to 2d (see FIG. 6). In this embodiment, ignition is performed in the order of the first cylinder 2a → the third cylinder 2c → the fourth cylinder 2d → the second cylinder 2b, and the exhaust stroke and the like are performed in this order.

  The piston 5 is connected to a crankshaft 7 via a connecting rod 8, and the crankshaft 7 rotates around a central axis in accordance with the reciprocating motion of the piston 5.

  A combustion chamber 6 is formed above the piston 5. An intake port 9 and an exhaust port 10 are opened in the combustion chamber 6, and an intake valve 11 and an exhaust valve 12 that open and close the ports 9 and 10 are provided in the cylinder head 4. The illustrated engine is a so-called double overhead camshaft (DOHC) engine, and two intake ports 9 and two exhaust ports 10 are provided for each cylinder, and two intake valves 11 and two exhaust valves 12 are provided. It is provided one by one.

  Note that the “combustion chamber” narrowly refers to the space above the piston 5 at the top dead center, but the combustion chamber 6 in this specification refers to the upper side of the piston 5 regardless of the vertical position of the piston 5. It refers to the space formed in (combustion chamber in a broad sense).

  Here, the engine body 1 of the present embodiment is set so as to have a relatively high geometric compression ratio of 14 or more for the purpose of improving the theoretical thermal efficiency, stabilizing the compression auto-ignition combustion described later, and the like. Yes. Since the upper limit value of the geometric compression ratio is considered to be about 20 from a practical viewpoint, the geometric compression ratio of the engine body 1 is set to an appropriate value in the range of 14 to 20. Is set.

  Various sensors are attached to the engine. For example, a water temperature sensor SW1 for detecting the temperature of the engine coolant, a rotation angle (crank angle) of the crankshaft 7, and a crank angle sensor SW2 for detecting the rotation speed of the engine, and the camshaft angle are detected. Cam angle sensor SW3 that outputs a signal for cylinder discrimination (determination of whether each cylinder is in the intake, compression, expansion, or exhaust stroke), outside air temperature for detecting the temperature of fresh air flowing into the combustion chamber A sensor SW4 is attached to the engine body.

  The intake valve 11 and the exhaust valve 12 are driven to open and close in conjunction with the rotation of the crankshaft 7 by valve mechanisms 13 and 14 including a pair of camshafts (not shown) disposed in the cylinder head 4. The

  CVVL 15 and VVT 16 are respectively incorporated in the valve operating mechanism 13 for the intake valve 11. The CVVL 15 is called a continuously variable valve lift mechanism and continuously (steplessly) changes the lift amount of the intake valve 11. The VVT 16 is called a variable valve timing mechanism, and variably sets the opening / closing timing (phase angle) of the intake valve 11. These CVVL15 and VVT16 are provided so that the lift amount and opening / closing timing of all the intake valves 11 of the engine can be changed. When both CVVL15 and VVT16 are driven, a pair of intake valves 11 in each cylinder 2 is provided. The lift amount and the opening / closing timing are changed simultaneously.

  The CVVL 15 configured as described above is already known. As a specific example thereof, a link mechanism that reciprocally swings the cam for driving the intake valve 11 in conjunction with the rotation of the camshaft, and the arrangement of the link mechanism (lever ratio). ) Variably set, and a stepping motor that changes the swing amount of the cam (the amount by which the intake valve 11 is pushed down) by electrically driving the control arm. For example, refer to JP 2007-85241 A). Also, as the VVT 16, various types such as a hydraulic type, an electromagnetic type, and a mechanical type are already known, and an appropriate one can be adopted.

  The valve mechanism 14 for the exhaust valve 12 incorporates a VVT 17 that is a variable valve timing mechanism that variably sets the opening / closing timing (phase angle) of the exhaust valve 12. Hereinafter, the CVVL 15 is referred to as an intake CVVL 15, the VVT 16 for the intake valve 11 is referred to as an intake VVT 16, and the VVT 17 for the exhaust valve 12 is referred to as an exhaust VVT 17.

  The cylinder head 4 of the engine body 1 is provided with one set of spark plugs 20 and injectors 21 for each cylinder 2.

  The injector 21 is provided so as to face the combustion chamber 6 from the ceiling surface (the lower surface of the cylinder head 4 covering the combustion chamber 6). A fuel supply pipe 23 is connected to the injector 21 of each cylinder 2, and fuel (fuel mainly composed of gasoline) supplied through each fuel supply pipe 23 is injected from the tip of the injector 21. .

  More specifically, on the upstream side of the fuel supply pipe 23, a high pressure fuel pump comprising a plunger type pump or the like linked to the crankshaft 7 is connected, and the high pressure fuel pump and the fuel supply are connected. A common rail for pressure accumulation common to all the cylinders is provided between the pipe 23. Then, the fuel accumulated in the common rail is supplied to the injectors 21 of the respective cylinders 2, whereby high pressure fuel of 30 MPa or more is injected from the injectors 21. Since the upper limit value of the fuel injection pressure is considered to be about 120 MPa from a practical viewpoint, the injection pressure from the injector 21 is set to an appropriate value in the range of 30 MPa to 120 MPa.

  The injector 21 is a so-called multi-hole injector, and has twelve nozzle holes at the tip. The installation portions of these injection holes (tip portions of the injectors 21) are located in the center of the ceiling of the combustion chamber 6, and each injection hole is perforated so that the opening direction faces obliquely downward on the outer side in the bore radial direction. Yes. For this reason, when fuel is injected from each injection hole of the injector 21, the fuel is injected radially so as to expand outward in the bore radial direction as it approaches the crown surface (upper surface) of the piston 5.

  The spark plug 20 is disposed adjacent to the injector 21 so as to face the combustion chamber 6 of each cylinder 2 from above. Specifically, the spark plug 20 has an electrode exposed to the combustion chamber 6 at the tip, and discharges a spark from the electrode in response to power supply from an ignition circuit (not shown).

  An intake passage 28 and an exhaust manifold 50 are connected to the intake port 9 and the exhaust port 10 of the engine body 1, respectively. A catalyst device 60 is connected to the downstream side of the exhaust manifold 50. Intake air (fresh air) from the outside is supplied to the combustion chamber 6 through the intake passage 28, and exhaust gas (burned gas) generated in the combustion chamber 6 passes through the exhaust manifold 50 and the catalyst device 60. It is discharged outside. Detailed structures of the exhaust manifold 50 and the catalyst device 60 will be described later. Hereinafter, the exhaust manifold 50 and the catalyst device 60 may be collectively referred to as an exhaust system.

  The intake passage 28 is provided with a common passage portion 28c formed of a single passage, a surge tank 28b provided at a downstream end portion of the common passage portion 28c, and a branch for each cylinder 2, and the surge tank 28b. And a branch passage portion 28 a that connects the intake port 9 of each cylinder 2.

  An external EGR device 30 is provided between the intake passage 28 and the exhaust system to recirculate a part of the exhaust gas passing through the exhaust system to the intake passage 28. Specifically, the external EGR device 30 is provided in an EGR passage 31 that connects each common passage portion 28 c of the intake passage 28 and an upstream end 61 of a casing 62 described later of the catalyst device 60, and in the middle of the EGR passage 31. The EGR valve 32 controls the flow rate of the exhaust gas passing through the interior, and the water-cooled EGR cooler 33 that cools the temperature of the exhaust gas passing through the EGR passage 31.

  In the common passage portion 28c of the intake passage 28, a throttle valve 25 is provided so as to be opened and closed. However, in this embodiment, the lift amount of the intake valve 11 is adjusted by the intake CVVL 15, and the opening / closing timing of the exhaust valve 12 is adjusted by the exhaust VVT 17, so that the EGR gas amount and fresh air in the combustion chamber 6 are adjusted. Since the amount is adjusted, the throttle valve 25 is kept fully open except when the engine is stopped.

  FIG. 3 is an enlarged view for specifically explaining the shape of the crown surface of the piston 5. As shown in FIG. 3 and FIG. 1 above, a concave cavity 40 is provided at the center of the crown surface of the piston 5. The cavity 40 has an upward opening 40a facing the injector 21 at the upper end, and the area (opening area) of the opening 40a is the maximum cross-sectional area inside the cavity 40 (each height of the cavity 40). The maximum horizontal cross-sectional area at the position) is set. That is, the cavity 40 is formed in a constricted shape so that the inner diameter becomes narrower toward the upper side in the range from the opening 40a to a predetermined depth.

(2) Configuration of Exhaust System FIG. 4 is a side view of a part of FIG. In FIG. 2, an outer tube 58 described later is cut and the inside of the outer tube 58 is exposed so that the configuration becomes clearer. As shown in FIGS. 4 and 2, the exhaust manifold 50 includes, in order from the upstream side, three independent exhaust passages 52, a substantially cylindrical assembly portion 56, and a substantially cylindrical connection portion 57. Yes.

  Each independent exhaust passage 52 is connected to the exhaust port 10 of each cylinder 2. Specifically, among the cylinders 2, the exhaust port 10 of the first cylinder 2a and the exhaust port 10 of the fourth cylinder 2d are individually connected to independent exhaust passages 52a and 52d, respectively. On the other hand, the exhaust ports 10 of the second cylinder 3b and the third cylinder 2c, whose exhaust strokes are not adjacent to each other and the exhaust order is not continuous, are not exhausted from these cylinders 2b, 2c at the same time, so the structure is simplified. From this point of view, it is connected to one independent exhaust passage 52b. More specifically, the independent exhaust passage 52b connected to the exhaust ports 10 of the second cylinder 2b and the third cylinder 2c is separated into two passages on the upstream side, and the second cylinder 2b is connected to one of them. The exhaust port 10 of the third cylinder 2c is connected to the other exhaust port 10. In the present embodiment, the independent exhaust passage 52 corresponding to the exhaust port 10 of the second cylinder 2b and the third cylinder 2c is linearly opposed to the central portion of the cylinders 2b and 2c, that is, the substantially central portion of the engine body 1. The independent exhaust passages 52 extending and corresponding to the exhaust ports 10 of the other cylinders 12a and 12d are independent exhaust passages corresponding to the second cylinder 3b and the third cylinder 2c from positions corresponding to the corresponding exhaust ports 10. Curved and extended toward 52.

  These independent exhaust passages 52 are independent from each other, and the exhaust discharged from the second cylinder 2b or the third cylinder 2c, the exhaust discharged from the first cylinder 2a, and the exhaust discharged from the fourth cylinder 2d Are discharged downstream through the independent exhaust passages 52 independently of each other. The gas that has passed through each independent exhaust passage 52 flows into the collecting portion 56.

  Each of the independent exhaust passages 52 and the collective portions 56 are arranged around the high-speed exhaust as the exhaust is ejected from the independent exhaust passages 52 at a high speed and the exhaust flows into the collective portion 56 at a high speed. Due to the generated negative pressure action, that is, the ejector effect, negative pressure is generated in another adjacent independent exhaust passage 52 and the exhaust port 10 communicating with the independent exhaust passage 52, and the gas in the exhaust port 10 is sucked out downstream. It has such a shape.

  Specifically, each of the independent exhaust passages 52 has a shape in which the flow passage area becomes smaller toward the downstream so that the exhaust is ejected from the independent exhaust passages 52 into the collecting pipe 56 at a high speed. Yes. In the present embodiment, as shown in FIG. 5, each independent exhaust passage 52 is reduced in cross-sectional area from the upstream portion having a substantially elliptical cross section toward the downstream, and at the downstream end of the upstream portion. It has a sector shape that is approximately 1/3 of the elliptical cross-sectional area. The independent exhaust passages 52 are aggregated and connected to the collective portion 56 so that their downstream ends forming a sector shape are adjacent to each other to form a substantially circular cross section as a whole.

  The collecting portion 56 has a shape in which the flow passage area becomes smaller toward the downstream side so that the exhaust discharged from each of the independent exhaust passages 52 flows downstream while maintaining a high speed. In the present embodiment, the flow passage area at the downstream end of the collecting portion 56 is set to be smaller than the total flow passage area at the downstream end of each independent exhaust passage 52 in order to further increase the exhaust speed.

  Here, the cross-sectional area of the downstream end of the independent exhaust passage 52 is the same as the flow area of the downstream end of the gathering portion 56, where a (see FIG. 4) is the diameter of a perfect circle having the same area as this cross-sectional area. Assuming that the diameter of a perfect circle having an area is D (see FIG. 4), if a / D is set in the range of a / D ≧ 0.5, the exhaust at the collecting portion 56 is sufficiently high. It has been found that a high ejector effect can be obtained by passing. Therefore, in the present embodiment, the independent exhaust passage 52 and the like are configured to satisfy a / D ≧ 0.5 in addition to the above configuration. In addition, in order to further increase the inflow speed of the exhaust gas from the independent exhaust passage 52 to the collecting portion 56, a portion having a reduced flow area, that is, a throttle portion is provided at the downstream end of the independent exhaust passage 52. Preferably, the diameter of the flow passage area of the throttle portion is a, and the connecting portion 57 is shaped so that a / D ≧ 0.5.

  Further, in the present embodiment, in order to prevent pressure nonuniformity from occurring in the collecting portion 56 and the connecting portion 57 and thereby reducing the gas suction force to the downstream side, the upstream portion of the collecting portion 56 is suppressed. The inner surface is separated from the downstream end of the independent exhaust passage 52 in a direction orthogonal to the gas flow direction. Specifically, the upstream portion 56a of the collecting portion 56 has a larger inner diameter than the cylindrical portion formed at the downstream end of each of the independent exhaust passages 52 and has a substantially constant flow path area in the upstream / downstream direction. . And the downstream part 56b of the gathering part 56 is made into the shape which a flow-path area reduces as it goes downstream.

  The connecting portion 57 constitutes an upstream portion thereof, and constitutes a straight portion 57a having a substantially constant channel area in the upstream and downstream directions and a downstream portion thereof, is substantially frustoconical, and flows toward the downstream. It consists of a diffuser part 57b with an increased road area.

  In the collective portion 56 and the diffuser portion 57b configured as described above, the heat release amount of the exhaust gas is suppressed to be small, and the exhaust gas whose temperature is maintained high is discharged downstream.

  Specifically, as described above, the flow path area of the collecting portion 56, specifically, the downstream portion 56b of the collecting portion 56 is smaller toward the downstream side. Further, the flow passage area of the straight portion 57a of the connecting portion 57 is substantially constant. Therefore, the exhaust discharged from the independent exhaust passage 52 passes through the collecting portion 56 and the straight portion 57a at a high speed. At this time, since the pressure and temperature of the exhaust gas are reduced, the amount of heat released to the outside of the exhaust gas can be kept small in the collecting portion 56 and the straight portion 57a. Then, the exhaust gas that has passed through the straight portion 57a flows into the diffuser portion 57b whose flow area increases as it goes downstream, so that its pressure and temperature are recovered and discharged to the downstream side while maintaining a high temperature. The

  Further, in the present embodiment, the collecting portion 56 and the connecting portion 57 are inserted into a hollow outer tube 58, and the portion from the upstream end of the collecting portion 56 to the downstream end of the connecting portion 57 is a double tube. It has a structure. Therefore, when passing through the collecting portion 56 and the connecting portion 57, heat radiation to the outside of the exhaust is further suppressed, and exhaust at a higher temperature is discharged downstream.

  The catalyst device 60 is a device for purifying the exhaust discharged from the engine body 1. The catalyst device 60 includes a catalyst body (catalyst) 64 and a casing 62 that houses the catalyst body 64. The casing 62 has a substantially cylindrical shape extending in parallel with the exhaust flow direction. The catalyst body 64 is for purifying harmful components in the exhaust gas, and has a three-way catalyst function in an atmosphere having a theoretical air-fuel ratio. The catalyst body 64 contains, for example, a three-way catalyst.

  The catalyst body 64 is accommodated in a central portion in the upstream and downstream direction of the casing 62, and a predetermined space is formed at the upstream end 61 of the casing 62. The downstream end of the diffuser portion 57b is connected to the upstream end 61 of the casing 62, and the exhaust gas discharged from the diffuser portion 57b flows into the upstream end 61 of the casing 62 and then proceeds to the catalyst body 64 side. The EGR passage 31 is connected to the upstream end 61 of the casing 62, and a part of the exhaust discharged from the diffuser portion 57b flows into the EGR passage 31 according to the opening / closing operation of the EGR valve 32.

(4) Effects of the exhaust system In the exhaust system configured as described above, exhaust gas is ejected and flows from the independent exhaust passage 52 to the collecting portion 56 and the connecting portion 57 at a high speed, and in response to this, other independent effects are obtained by the ejector effect. A negative pressure is generated in the exhaust passage 52. Therefore, as will be described later, when the exhaust valve 12 of the predetermined cylinder (exhaust stroke cylinder) 2 is opened in the low-speed and high-load region, another independent order in which the exhaust sequence is set immediately before the exhaust stroke cylinder 2 is established. By opening the intake valve 11 and the exhaust valve 12 of the cylinder (intake stroke cylinder) 2 communicating with the exhaust passage 52 to overlap each other, the exhaust gas discharged from the exhaust stroke cylinder 2 is caused to overlap. The generated negative pressure can be applied to the intake stroke cylinder 2 to promote scavenging of the intake stroke cylinder 2, avoiding occurrence of abnormal combustion such as knocking that is likely to occur in a low speed and high load region under a high compression ratio, and combustion Promotion of introduction of fresh air into the chamber 6 can be realized. That is, it is possible to realize a high compression ratio of the engine while ensuring a high engine torque.

  FIG. 7 shows the results of examining the volumetric efficiency and knock limit torque (maximum engine torque in a range where knocking does not occur) at a full load of 1500 rpm. In FIG. 7, the solid line is the result in the exhaust system of the present engine, and the broken line is the result in the so-called 4-2-1 type exhaust system in which the exhaust passage is lengthened to promote scavenging. These are the results for the so-called 4-1 type exhaust system. Note that the 4-1 type exhaust system is an exhaust system that has been widely used in the past that does not particularly consider scavenging performance, and the four exhaust passages respectively connected to the exhaust port 10 are joined in the middle. There is no one.

  As shown in FIG. 7, in the exhaust system of the engine, the maximum volume efficiency can be increased compared to other exhaust systems, and scavenging is further promoted (the residual gas amount in the combustion chamber 6 is further increased). Therefore, it is not necessary to adjust the combustion timing to the retard side in order to avoid knocking, and the engine torque can be increased even in the same volumetric efficiency.

  Moreover, the exhaust system of this engine promotes scavenging by utilizing the ejector effect, and the so-called 4-2-1 type exhaust in which the exhaust passage is lengthened to promote scavenging while ensuring high engine torque. The distance L (see FIG. 2) from the exhaust port 10 (specifically, the opening portion of the cylinder head 4 where the exhaust port 10 opens) to the catalytic device 60 (specifically, the upstream end 61 of the casing 62 of the catalytic device 60) rather than the system. ) Can be shortened. Therefore, the temperature of the exhaust gas flowing into the catalyst main body 64 can be increased, and the activity of the catalyst main body 64 can be promoted. For example, in the 4-2-1 type exhaust system, the distance L from the exhaust port 10 to the catalyst device 60 is about 600 mm, whereas in the exhaust system of the present engine, this distance is maintained while the scavenging performance is similarly maintained. L can be shortened to about 400 mm.

  Further, as described above, in the exhaust system of this engine, the exhaust discharged from the independent exhaust passage 52 passes through the collecting portion 56 and the connecting portion 57 at a high speed, and this exhaust passes around the other independent exhaust passage 52. In addition to being suppressed from expanding and lowering in temperature, the pressure and temperature of the exhaust gas are recovered by the diffuser portion 57b of the connecting portion 57, and the heat radiation at the collecting portion 56 and the connecting portion 57 is suppressed to be small. The temperature flowing into the catalyst body 64 can be increased. Furthermore, since the part from the upstream end of the said gathering part 56 to the downstream end of the connection part 57 is made into the double pipe structure, the temperature which flows in into the catalyst main body 64 can be made high.

(5) Control System FIG. 8 is a block diagram showing an engine control system. The ECU 100 shown in FIG. 8 is a device (control means according to the present invention) for comprehensively controlling each part of the engine, and includes a well-known CPU, ROM, RAM, and the like.

  Various information is input to the ECU 100 from various sensors provided in the engine. That is, the ECU 100 is electrically connected to the water temperature sensor SW1, the crank angle sensor SW2, the cam angle sensor SW3, the outside air temperature sensor SW4, and the like provided in the engine, and input signals from these sensors SW1 to SW4. Based on the above, various information such as engine coolant temperature, crank angle, engine speed, cylinder discrimination information, and fresh air temperature are acquired.

  In addition, information from various sensors provided in the vehicle is also input to the ECU 100. For example, the vehicle is provided with an accelerator opening sensor SW5 that detects an opening degree of an accelerator pedal (accelerator opening degree) that is not shown in the figure that is depressed by the driver, and the accelerator opening degree sensor SW5 detects the accelerator opening degree sensor SW5. The accelerator opening is input to the ECU 100.

  A more specific function of the ECU 100 will be described. The ECU 100 includes, as main functional elements, a determination unit 101, an injector control unit 102, an intake control unit 103, an external EGR control unit 105, and an ignition control unit 106. have.

  The determination means 101 determines the engine based on the engine speed specified from the detected value of the crank angle sensor SW2 and the engine load (target torque) specified from the detected value of the accelerator opening sensor SW4. It is determined each time whether it should be controlled in a manner. In the following, it is assumed that the engine rotation speed is Ne and the engine load is T.

  FIG. 9 is a setting diagram (control map) showing the types of control determined based on the engine rotational speed Ne and the load T. During operation of the engine, the determination unit 101 determines the control content of the engine so as to follow the control map of FIG. Note that the control map in FIG. 9 is basically for a warm state where the coolant temperature detected by the engine coolant temperature sensor SW1 is equal to or higher than a predetermined value (for example, 80 ° C.). Although the control map when the engine is in the cold state may be different from that in FIG. 9, the description thereof is omitted here.

  In the control map of FIG. 9, in the region where the engine load T is low (low load region), the first operation region A1 is set over all engine rotational speed regions including the low speed region (region where the engine rotational speed is low). In the high load region where the load T is higher than the first operation region A1, the second operation region A2 is set over all engine rotation speed regions including the low speed region (region where the engine rotation speed is low). .

  During the operation of the engine, which operating region (A1, A2) in FIG. 9 corresponds to the operating point of the engine (point on the control map specified from each value of the load T and the rotational speed Ne). Is determined each time, and appropriate control corresponding to each operation region is executed.

  Referring back to FIG. 8 again, the injector control means 102 electromagnetically opens and closes a needle valve (not shown) incorporated in the injector 21 (a valve for opening and closing the nozzle at the tip of the injector 21). The injection quantity and injection timing of the fuel injected into the combustion chamber 6 are controlled.

  The intake air control means 103 drives the intake air CVVL 15 and the intake air VVT 16 to change the lift amount (opening amount) and the opening / closing timing of the intake valve 11, and drives the exhaust air VVT 17 to open / close the exhaust valve 12. Control to change the timing is performed.

  The external EGR control means 105 switches the presence / absence of an operation (external EGR) for returning the exhaust gas from the exhaust passage 29 to the intake passage 28 by adjusting the opening degree of the EGR valve 32 provided in the EGR passage 31. At the same time, the exhaust gas recirculation amount (external EGR amount) by the external EGR is controlled.

  The ignition control means 106 controls the timing of spark ignition (ignition timing) by the spark plug 20 and the like.

(6) Control contents in each operation region Next, based on the control of the ECU 100 having the above functions, what kind of control is performed in each operation region (A1, A2) shown in FIG. I will explain. As a premise of the description here, it is assumed that the engine coolant temperature is sufficiently warm (that is, the operation is warm). When the engine is thus operated in a warm state, the ECU 100 determines the engine operating point (load T and rotational speed Ne) based on the detected values of the crank angle sensor SW2 and the accelerator opening sensor SW4. It is sequentially determined which operation region corresponds to the control map of FIG. Then, depending on whether the determined operation region is the first operation region A1 or the second operation region A2 in FIG. 9, the following control is executed.

(I) 1st operation area A1
In the first operation region A1, the ignition of the air-fuel mixture by the spark plug 20 is stopped, and the mixture of fuel and air that has been sufficiently mixed is self-ignited by the compression action of the piston 5 (Homogenous-Charge Compression). An ignition combination (premixed compression auto-ignition combustion) mode is executed.

  FIG. 10 is a diagram showing the fuel injection timing, the lift characteristics of the intake and exhaust valves 11 and 12, and the heat generation rate (J / deg) generated by combustion based on the fuel injection timing when the lean HCCI combustion mode is executed. As shown in FIG. 10, in the lean HCCI combustion mode, fuel is injected at a timing sufficiently earlier than the compression top dead center (TDC between the compression stroke and the expansion stroke), and the fuel and air are sufficiently mixed. The air-fuel mixture is ignited by the compression action of the piston 5 in the vicinity of the mixed compression top dead center.

  Specifically, in this embodiment, when operating in this lean HCCI combustion mode, a relatively small amount of fuel is injected (P) from the injector 21 into the combustion chamber 6 at a predetermined time during the intake stroke. The homogeneous and lean air-fuel mixture formed based on a small amount of fuel collectively injected by the fuel injection P and air (new air) introduced from the intake passage 28 into the combustion chamber 6 is compressed by the piston 5. Due to the action, the temperature and pressure are increased, and self-ignition occurs near the compression top dead center. Then, based on such self-ignition, combustion accompanied by heat generation as shown by the waveform Qa occurs.

  In the lean HCCI combustion mode, G / F, which is the ratio of the weight G of the total gas in the combustion chamber 6 to the fuel weight F injected into the combustion chamber 6, is set to be 30 or more (for example, 35). The However, under such a lean and homogeneous air-fuel ratio, misfire may occur unless the temperature in the combustion chamber is intentionally increased. As described above, in this gasoline engine, the geometric compression ratio of the engine body is set to a high value of 14 or higher, and the temperature of the combustion chamber can be raised to a certain extent by the compression action of the piston 5, In order to realize more stable combustion in a low load region where the temperature in the room is low, in the present embodiment, control for securing a large amount of internal EGR gas in the combustion chamber 6 is performed. Specifically, in the lean HCCI combustion mode, as shown in the lift curve EX (lift curve of the exhaust valve 12) and the lift curve IN (lift curve of the intake valve 12) in FIG. Are not simultaneously opened, and the intake CVVL15, the intake VVT16, and the exhaust VVT17 are the intake valve 11 and the exhaust valve 12 so that a negative overlap period is formed in which both the exhaust valve 12 and the intake valve 11 are closed. Thus, most of the exhaust gas generated in the combustion chamber 6 remains in the combustion chamber 6 (internal EGR is performed). Thus, by leaving the high-temperature exhaust gas (internal EGR gas) in the combustion chamber 6, the combustion chamber 6 is heated to a high temperature and stable compression auto-ignition combustion is realized.

  Here, when the load T is high and the fuel injection amount is increased, the required fresh air amount is increased. Therefore, in this lean HCCI combustion mode, the amount of internal EGR gas is decreased as the load T increases, and the introduction of fresh air is promoted. Specifically, as the load T increases, the closing timing of the exhaust valve 12 is changed to the retard side, such as from the solid lift curve EX to the broken lift curve EX in FIG. The amount of exhaust gas discharged to the independent exhaust passage 52 side is increased. Further, the lift amount of the intake valve 11 is gradually increased as the load T increases. 10 is a lift curve when the intake valve 11 is in a small lift state, and when the load T increases from this state, the lift amount of the intake valve 11 is accordingly a lift curve with a broken line. The upper limit is gradually increased. Thus, when the lift amount of the intake valve 11 is increased, the closing timing of the intake valve 11 is fixed in the vicinity of the intake bottom dead center (BDC between the intake stroke and the compression stroke) while the intake valve 11 is closed. The opening / closing timing and lift amount of the intake valve 11 are adjusted by the intake CVVL 15 and the intake VVT 16 so that only the opening timing is gradually advanced toward the exhaust top dead center (TDC between the exhaust stroke and the intake stroke). .

  In this lean HCCI combustion mode, the opening degree of the EGR valve 32 provided in the EGR passage 31 is fully closed, and the exhaust gas recirculation from the exhaust passage 29 to the intake passage 28, that is, the external EGR is stopped.

  In this way, in the lean HCCI combustion mode, the lean air-fuel mixture having a G / F of 30 or more is self-ignited and combusted. Therefore, while suppressing the generation of NOx, the combustion temperature and thus the cooling loss are reduced, thereby reducing the thermal efficiency. (Fuel consumption) can be improved.

(Ii) Second operation area A2
In the second operation region A2 set in the high load region, since a large amount of fuel is injected, there is a problem in that combustion noise increases remarkably and knocking occurs when attempting to perform compression self-ignition combustion. is there. Therefore, in the second operation region A2, spark ignition combustion (SI combustion) is performed in which the air-fuel mixture is ignited and flame is propagated instead of compression self-ignition combustion.

  Here, also in SI combustion, when the temperature in the combustion chamber 6 is excessively high, knocking occurs. In particular, in this gasoline engine, the compression ratio is set to a very high value. Therefore, the temperature in the combustion chamber 6 tends to increase.

  Therefore, in the present gasoline engine, scavenging is promoted by utilizing the ejector effect in the second operation region A2, and the amount of residual gas (internal EGR gas) in the combustion chamber 6 is reduced to reduce the temperature in the combustion chamber 6. The engine torque is increased while avoiding knocking by securing new air volume and reducing the air volume.

  Specifically, in the second operation region A2, as shown in FIG. 6, the target valve timing of the intake valve 11 and the exhaust valve 12 is such that the opening period of the exhaust valve 12 and the opening period of the intake valve 11 are The exhaust valve 12 is adjusted so that it overlaps with the intake top dead center (TDC) interposed therebetween, and the exhaust valve 12 starts to open during the overlap period T_O / L of the other cylinders 12. Specifically, the exhaust valve 12 of the third cylinder 2c is opened during the period in which the intake valve 11 and the exhaust valve 12 of the first cylinder 2a overlap, and the intake valve 11 and the exhaust valve of the third cylinder 2c are opened. The exhaust valve 12 of the fourth cylinder 2d is opened during the period in which the second cylinder 2d overlaps with the second cylinder 2b during the period in which the intake valve 11 and the exhaust valve 12 of the fourth cylinder 2d overlap. The exhaust valve 12 of the first cylinder 2a is opened, and the exhaust valve 12 of the first cylinder 2a is adjusted to open during the period in which the intake valve 11 and the exhaust valve 12 of the second cylinder 2b overlap.

  As described above, in the second operation region A2, when the exhaust valve 12 of the predetermined cylinder (exhaust stroke cylinder) 2 is opened, another independent exhaust in which the exhaust order is set immediately before the exhaust stroke cylinder 2 is established. When the intake valve 11 and the exhaust valve 12 of the cylinder (intake stroke cylinder) 2 communicating with the passage 52 are overlapped, the exhaust valve 12 of the exhaust stroke cylinder 2 is opened and the exhaust stroke cylinder 2 is independently exhausted. As exhaust gas is discharged through the passage 52 to the collecting portion 56 at a high speed, a negative pressure is generated in the independent exhaust passage 52 of the intake stroke cylinder 2 during the overlap period and thus in the exhaust port 10 by the ejector effect. Scavenging is promoted. That is, a large amount of residual gas (internal EGR gas) in the combustion chamber 6 is discharged to the independent exhaust passage 52 side, and more fresh air is introduced into the combustion chamber 6, so that the engine is avoided while avoiding knocking. Torque can be increased. In particular, immediately after the opening of the exhaust valve 12 is started, exhaust (so-called blowdown gas) is discharged from the cylinder 12 at a very high speed, and accordingly, a high negative pressure is generated.

  In this engine, the valve opening timing and the valve closing timing of the intake valve 11 and the exhaust valve 12 are as follows. As shown in FIG. This is the time of lowering, for example, the time of 0.4 mm lift.

  Furthermore, in this embodiment, in order to more reliably avoid abnormal combustion such as knocking, fuel is injected considerably before the compression top dead center (for example, during the intake stroke), and spark ignition is performed near the compression top dead center. Instead of normal SI combustion, as shown in FIG. 11, fuel is injected from the injector 21 during the compression stroke (P1, P2), and spark ignition is performed on the spark plug 20 after the fuel injection P1, P2. Then, the rapid retarded SI combustion mode is executed in which the air-fuel mixture is combusted by flame propagation in a short time from the timing when the compression top dead center is passed (the initial stage of the expansion stroke). FIG. 11 is a diagram showing the fuel injection timing, the lift characteristics of the intake and exhaust valves 11 and 12, and the heat generation rate (J / deg) generated by combustion based on the fuel injection timing and the lift characteristics when the rapid retarded SI combustion mode is executed. . In this specification, the term “late stage” or “initial stage” of a certain process refers to the latter period or the initial stage when the process is divided into three stages: initial, middle, and late. For example, in the later stage of the compression stroke, it refers to the range of 60 to 0 ° CA before compression top dead center (BTDC), and in the early stage of the expansion stroke, it is from 0 to 60 ° CA after compression top dead center (ATDC). Will point to the range.

  Specifically, as shown in FIG. 11, in the rapid retarded SI combustion mode, fuel is injected from the injector 21 at a high pressure of 30 MPa or more divided into two injection timings (P1, P2) set in the latter half of the compression stroke. Is injected. As the timing of each fuel injection P1, P2, for example, the period from the start timing of the first injection P1 to the completion timing of the second injection P2 is approximately 20 ° to 0 ° before compression top dead center (BTDC). It is set to be within a period of about CA.

  In the rapid retarded SI combustion mode in which such injection control is performed, the injection period can be shortened by injecting fuel at a very high injection pressure of 30 MPa or more (for example, 40 MPa) as described above. At the same time, the fuel spray can be atomized, and a large amount of fuel can be sufficiently vaporized and atomized in a short time to form a relatively homogeneous (or weakly stratified) mixture. In addition, since the injection pressure is high, large turbulent energy can be maintained until the combustion chamber 6 passes a certain amount of compression top dead center at which the temperature and pressure of the combustion chamber 6 are the highest. Accordingly, combustion by spark ignition can be started for a short time after the fuel is injected, and thus within a period in which the attenuation of turbulent energy accompanying the fuel injection is small (in a period in which the turbulent energy is large). The combustion period can be shortened by large turbulent energy. As the combustion period is shortened, the air-fuel mixture can be burned out by proper flame propagation before combustion is caused. That is, high thermal efficiency and engine torque can be maintained while avoiding abnormal combustion such as knocking. In addition, since combustion starts at the timing when the compression top dead center is passed, the combustion temperature does not rise excessively, and combustion does not start with insufficient vaporization and atomization of fuel. The increase is avoided and the emission property is also maintained well.

  Further, the fuel is injected in two portions, and ignition is performed after the second fuel injection. The fuel is atomized by the first fuel injection P1, and the fuel is injected at the time of ignition by the second fuel injection. Turbulent energy can be increased. The turbulent energy is also increased by being ejected from as many as 12 nozzles. Further, as shown in FIG. 12, the combination of the multi-injector type injector 21 and the cavity 40 provided in the piston 5 prevents the fuel spray injected by the fuel injections P1 and P2 in the latter half of the compression stroke from being disturbed. Since the flow energy can be diffused mainly in the cavity 40, the combustion period can be further shortened.

  Here, if the ignition timing is further advanced than in the example of FIG. 11, the combustion start timing is closer to the compression top dead center, and further improvement in thermal efficiency and output torque can be expected. Since the knocking is likely to occur when it is advanced, the ignition timing is set to the advance side as much as possible under the restriction that knocking does not occur. Under such circumstances, the ignition timing is set, for example, within a range of about 0 ° to 20 ° CA after compression top dead center (ATDC).

  In the rapid retarded SI combustion mode, the new air-fuel ratio is set so that the average air-fuel ratio of the entire combustion chamber 6 becomes the stoichiometric air-fuel ratio (excess air ratio λ = 1) with respect to the total injection amount by the fuel injections P1 and P2. The volume is controlled. Specifically, in the second operation region A2, the CVVL 15 is driven according to the increase in the load T, and the lift amount of the intake valve 11 is increased (the lift curve IN in FIG. 11 is controlled from the chain line to the broken line side). Accordingly, fresh air corresponding to the fuel injection amount is introduced (lift curve IN in FIG. 11). In this embodiment, the lift amount is increased while the lift peak position of the intake valve 11 is fixed.

  In the rapid retarded SI combustion mode, external EGR is performed in which the exhaust gas is cooled by the EGR cooler 33 through the EGR passage 31 and then returned to the intake passage 28. Note that external EGR is prohibited in the vicinity of the full load of the engine in order to secure a larger amount of fresh air.

(7) Effects and the like As described above, in this engine, scavenging is promoted in the high load region A2 by the ejector effect, and the compression ratio of the cylinder 2 is set to a high compression ratio, so that the proper efficiency in the low load region A1 is obtained. While realizing compression self-ignition and improving thermal efficiency, knocking in a high load region can be avoided and high engine torque can be secured. In addition, since the scavenging is promoted by utilizing the ejector effect, the distance from the exhaust port 10 to the catalyst device 60 can be shortened as described above while maintaining high scavenging performance. The amount of heat released from the exhaust gas until reaching the main body 64 is kept small, and the temperature of the exhaust gas flowing into the catalyst main body 64 is increased, whereby the catalyst main body 64 is activated early and the active state of the catalyst main body 64 is maintained. Can be realized.

  In particular, in this engine, the exhaust discharged from the independent exhaust passage 52 passes through the collecting portion 56 and the connecting portion 57 at a high speed, and the exhaust gas temperature is recovered by the diffuser portion 57b of the connecting portion 57. It is possible to increase the temperature flowing into the catalyst body 64 by suppressing the expansion and lowering of the temperature by going around the other independent exhaust passage 52 and suppressing the heat radiation at the collecting portion 56 and the connecting portion 57. it can. Furthermore, since the part from the upstream end of the said gathering part 56 to the downstream end of the connection part 57 is made into the double pipe structure, the temperature which flows in into the catalyst main body 64 can be made high.

  Further, in this engine, the rapid retarded SI combustion mode is executed in the high load region A2, so that knocking can be avoided more reliably while securing the engine torque, and high thermal efficiency can be obtained in the entire operation region. Can do.

  Here, in the above-described embodiment, the case where the exhaust valve 12 and the intake valve 11 are negatively overlapped (not overlapped) in the first operation region A1 has been described, but the second operation region is also included in the first operation region A1. Similarly to A2, the exhaust valve 12 and the intake valve 11 may be overlapped. For example, in the region on the high load side in the first operation region A1, the compression self-ignition combustion can be realized even when the amount of internal EGR gas is relatively small due to the relatively high load. The exhaust valve 12 and the intake valve 11 may overlap in the high load side region of the operation region A1. Further, in the region on the high load side in the first operation region A1, the exhaust valve 12 of the cylinder 2 that is one exhaust stroke before the cylinder 2 is opened during the overlap period of the predetermined cylinder 2. You may do it. In this way, even in this operating region, scavenging can be promoted by the ejector effect to introduce more fresh air, and lean premixed compression self-ignition combustion can be realized up to a higher load region. .

  In the above embodiment, the exhaust valve 12 and the intake valve 11 are overlapped in the entire second operation region A2. However, in the operation region where the engine rotational speed is high and the exhaust flow rate is large, the back pressure is high. As a result of increase, sufficient ejector effect may not be obtained. Therefore, in such a high speed region, the exhaust valve 12 and the intake valve 11 are overlapped in order to use the ejector effect, and the exhaust valves 12 of the other cylinders 2 are opened during the overlap period. Control may be stopped. That is, the control may be performed such that the exhaust valve 12 and the intake valve 11 are negatively overlapped, or the overlap period and the opening period of the exhaust valves 12 of the other cylinders 2 are not overlapped.

  In the above embodiment, the case where the HCCI combustion mode is performed in the entire first operation region A1 has been described. However, in a region where the engine rotational speed is high, the heat absorption time of the fuel is reduced, so that proper compression self-ignition combustion is performed. May become difficult. Therefore, in such a case, the HCCI combustion mode is performed only in the low speed region of the first operation region A1, and other combustion modes (for example, SI combustion or self-ignition of the air-fuel mixture by spark ignition are performed in the high speed region. Combustion configured to be facilitated) may be performed.

  In the above embodiment, the case where the intake valve 11 and the exhaust valve 12 are negatively overlapped in order to increase the internal EGR gas amount in the first operation region A1 has been described. However, the exhaust valve 12 is added to the exhaust stroke. The exhaust valve 12 can be opened during the intake stroke, and the exhaust valve 12 is opened during the intake stroke in addition to the exhaust stroke, so that the hot exhaust gas discharged to the exhaust port 10 flows back into the combustion chamber 6. Thus, the EGR gas amount in the combustion chamber 6 may be increased. However, in this case, there is a possibility that the backflow of the exhaust gas into the combustion chamber 6 is not sufficiently performed due to the improvement of the scavenging performance by the ejector effect. Therefore, in this case, a bypass passage for bypassing at least a portion from the downstream portion of the independent exhaust passage 52 configured to obtain a high ejector effect to the downstream end of the connecting portion 57, and opening and closing the bypass passage. A control valve, and in the second operation region A2, the bypass valve is blocked by the control valve so as to obtain a high ejector effect, while in the first operation region A1, the bypass valve is opened by the control valve, It may be configured such that the ejector effect is not substantially exhibited, that is, the backflow of the exhaust gas into the combustion chamber 6 is prevented from being hindered by the ejector effect.

  In the above embodiment, the fuel injection timing and the ignition timing in various operation regions are illustrated using FIGS. 10, 11 and the like. However, these are merely examples, and the fuel injection timing and the ignition timing are It can be appropriately changed depending on engine characteristics and the like.

  Further, in the first operation region A1, which is the execution region of the lean HCCI combustion mode, a required amount of fuel is injected once (fuel injection P) during the intake stroke. Any timing that ensures a certain amount of fuel agitation time until compression top dead center (for example, before the middle of the compression stroke) may be used, and is not necessarily limited to during the intake stroke.

  In the above embodiment, the G / F is controlled to 30 or more in the lean HCCI combustion mode. However, the theoretical air-fuel ratio λ is controlled instead of G / F, and this theoretical air-fuel ratio is set to λ = 2 or more. You may implement control to.

  Further, in the second operation region A2, which is the execution region of the rapid retarded SI mode, the fuel injections P1 and P2 are performed in the latter half of the compression stroke, but the number and the injection timing are the latter half of the compression stroke. Within the range, it can be appropriately changed. For example, on the low load side of the second operation region A2 where the fuel injection amount is relatively small, collective injection that injects a required amount of fuel at a time may be performed. Conversely, on the high load side (near the maximum load) in the second operation region A2, fuel may be injected in three or more times.

  In the above embodiment, the external EGR for recirculating the exhaust gas from the exhaust passage 29 to the intake passage 28 is executed in the second operation region A2 that is the execution region of the rapid retarded SI mode. From the standpoint of achieving this, only the fuel and fresh air corresponding to λ = 1 are filled in the combustion chamber 6 by stopping the external EGR and adjusting the flow rate (throttling) by the throttle valve 25. Good.

  On the other hand, when the external EGR that introduces the low-temperature exhaust gas sufficiently cooled in the EGR cooler 33 to the intake side in the rapid retarded SI mode as in the above-described embodiment, abnormalities such as pre-ignition and knocking are performed. Combustion can be prevented more effectively. For this reason, in an engine that has a particularly high compression ratio and is prone to abnormal combustion, it is preferable to execute an external EGR that introduces this low-temperature exhaust gas when implementing the rapid retarded SI combustion mode from the viewpoint of improving exhaust performance. .

  Further, instead of the rapid retarded SI combustion mode in the high load region, normal SI combustion may be performed in which fuel is injected well before the compression top dead center and spark ignition is performed near the compression top dead center. However, as described above, if the rapid retarded SI combustion mode is performed, the thermal efficiency can be improved while more reliably avoiding abnormal combustion.

  In the above embodiment, the injector 21 is a multi-hole injector, and twelve nozzle holes are provided at the tip thereof. However, the number of nozzle holes is not limited to twelve and may be more than twelve. It may be less. However, if the number of injection holes is too small, the concentration of the fuel injected from the injector 21 varies greatly in the circumferential direction. For this reason, it is desirable that the number of nozzle holes be eight or more.

  Moreover, in the said embodiment, although the injection pressure of the fuel from the injector 21 was made constant at 30 Mpa or more, the injection pressure does not need to be constant and is variably set according to the operation region. Also good. However, even in this case, at least in the execution region (second operation region A2) of the rapid retard SI mode, the fuel injection pressure is set to 30 MPa or more.

2 cylinders 6 combustion chambers 10 exhaust ports 11 intake valves 12 exhaust valves 21 injectors 52 independent exhaust passages 57 collecting portions 60 catalyst devices 64 catalyst bodies (catalysts)
100 ECU (control means)

Claims (7)

A plurality of cylinders each having a combustion chamber, an intake port, an exhaust port, an intake valve capable of opening and closing the intake port, and an exhaust valve capable of opening and closing the exhaust port; A gasoline engine provided with an injector for injecting fuel into the combustion chamber,
Control means for controlling the operation of the intake and exhaust valves and the combustion mode in the combustion chamber;
An independent exhaust passage connected to exhaust ports of one cylinder or a plurality of cylinders whose exhaust sequences are not continuous with each other;
A collecting portion connected to the downstream end of each independent exhaust passage so that the gas that has passed through each independent exhaust passage gathers;
A gas having a three-way catalyst function under an atmosphere of a theoretical air-fuel ratio, arranged downstream of the collecting portion so that the gas that has passed through the collecting portion flows in,
The geometric compression ratio of the cylinder is set to 14 or more,
The independent exhaust passages connected to the cylinders in which the exhaust order is continuous among the independent exhaust passages are connected to the collecting portion at positions adjacent to each other,
The downstream end portion of each independent exhaust passage is connected to another independent exhaust passage adjacent to each other by an ejector effect as exhaust is discharged from each cylinder through each exhaust port and each independent exhaust passage to the collective portion. The downstream side has a shape in which the flow area is smaller than the upstream side so that a negative pressure is generated in the exhaust port connected to the exhaust passage,
The control means executes a self-ignition combustion mode in which the air-fuel mixture in the combustion chamber burns by self-ignition at least in a low load and low speed region of the engine, and at least in a high load and low speed region of the engine. Between the cylinders in which the intake valve opening period and the exhaust valve opening period overlap with each other by a predetermined overlap period and the exhaust sequence is continued, the overlap period of one cylinder is opened in the other cylinder. A gasoline engine characterized in that an intake valve and an exhaust valve of each cylinder are opened so as to overlap with a valve timing. The gasoline engine according to claim 1,
The gasoline engine according to claim 1, wherein the collecting portion has a shape in which a flow path area is smaller on the downstream side than on the upstream side. The gasoline engine according to claim 2, wherein
A gasoline engine having a connecting portion disposed between the collecting portion and the catalyst and having a shape in which a flow path area is larger on the downstream side than on the upstream side. The gasoline engine according to any one of claims 1 to 3,
The control means is configured to open the intake valve opening period of each cylinder in at least a part of the low load and low speed range of the engine that executes the auto-ignition combustion mode in addition to the high load and low speed range of the engine. And the opening period of the exhaust valve overlap with each other by a predetermined overlap period, and the overlap period of one cylinder is between the cylinders in which the exhaust order is continuous at the time when the exhaust valve of the other cylinder is open A gasoline engine characterized in that an intake valve and an exhaust valve of each cylinder are opened so as to overlap. In the gasoline engine according to any one of claims 1 to 4,
The control means can control G / F, which is a ratio of the weight G of the total gas in the combustion chamber to the fuel weight F injected into the combustion chamber by the injector, and executes the self-ignition combustion mode. In a low load and low speed region of the engine, the ratio G / F of the total gas weight G in the combustion chamber to the fuel weight F is controlled to 30 or more, and the total gas weight G in the combustion chamber with respect to the fuel weight F toward the higher load side. A gasoline engine characterized by reducing the ratio G / F. In the gasoline engine according to any one of claims 1 to 5,
The control means can control the operation of the injector, and in the low load and low speed range of the engine that executes the auto-ignition combustion mode, the injector causes fuel to enter the combustion chamber in the first half of the intake stroke or the compression stroke. A gasoline engine characterized by being injected. In the gasoline engine according to any one of claims 1 to 6,
An ignition plug for supplying ignition energy to the air-fuel mixture in the combustion chamber;
The control means is capable of controlling the operation of the spark plug, and in a region on a higher load side than an operation region in which the auto-ignition combustion mode is executed, an average air in the combustion chamber is higher than that in the auto-ignition combustion mode. By executing fuel injection for injecting a large amount of fuel that makes the fuel ratio rich from the injector at an injection pressure of 30 MPa or more and spark ignition by the spark plug within a period from the late stage of the compression stroke to the initial stage of the expansion stroke, A gasoline engine that executes a rapid retarded SI mode in which an air-fuel mixture based on the combustion injection is rapidly combusted by flame propagation after a compression top dead center has passed a predetermined period or more.

Images