JP4862592B2 - Spark ignition gasoline engine - Google Patents

Description

  The present invention relates to a spark ignition gasoline engine.

In general, as disclosed in Non-Patent Document 1, a spark-ignition gasoline engine theoretically follows an Otto Cycle, and its theoretical thermal efficiency is η _th.
η _th = 1− (1 / ε ^κ−1 ) (1)
(Where ε is a compression ratio and κ is a specific heat ratio).

  As is clear from the equation (1), the theoretical thermal efficiency of the spark-ignition gasoline engine (therefore, shown in the figure, net thermal efficiency) is improved to a certain level when the compression ratio is high. In this regard, Non-Patent Document 1 discloses theoretical thermal efficiency by various compression ratios (8 ≦ ε ≦ 20) when a spark ignition gasoline engine is operated at 2000 rpm in a throttle fully open state (so-called WOT: Wide-Open Throttle). The research which investigated the change of is introduced. According to the description, the theoretical thermal efficiency and mean effective pressure (MEP) increase proportionally until the compression ratio reaches around 17, and then remain flat.

  Against the background of the above research results, practical application of a high compression engine has been attempted.

  However, in a spark-ignition engine with a high compression ratio, output reduction due to knocking in a high load operation region including the full load operation region is unavoidable.

  In this regard, as a general countermeasure against knocking, ignition retard that retards the ignition timing is widely known. However, in the high load operation region including the full load operation region, it has been considered that knocking avoidance by ignition retard greatly reduces the output and greatly impairs the merchantability.

  FIG. 1 is a graph showing an example of ignition retard during high load operation.

  For example, as shown in FIG. 1, with a compression ratio (ε = 11) widely used in ordinary engines, knocking does not occur when the ignition timing is set to 4 ° before compression top dead center, but a high compression ratio ( In the case of ε = 13), knocking occurs even if the ignition timing is 4 ° before compression top dead center. Therefore, it has been considered that a large ignition timing retard is necessary to employ a high compression ratio. This means that when the compression ratio is increased to about 13, the output reduction due to the retard of the ignition timing for preventing knocking surpasses the output increase due to the compression ratio improvement, and the output is greatly reduced. For this reason, conventionally, in consideration of the output reduction due to the ignition timing retard, in the high load operation region including the full load operation region, the compression ratio 12 is set as the limit of the high compression ratio, and a high compression ratio higher than that is set. I didn't use it.

  In view of this, for a high load operation region including a full load operation region, a method of lowering the effective compression ratio using a so-called Atkinson cycle or a Miller cycle is known. However, if the closing timing of the intake valve is changed during high load operation and the effective compression ratio is lowered, the fresh air is lost and the pressure is reduced and the charging efficiency is lowered and the output is reduced in the intake stroke.

Thus, a technique for reducing the actual compression ratio of the engine in a high load operation region including the full load operation region is also known. For example, Patent Documents 1 and 2 disclose a technique in which a variable compression ratio mechanism for changing an actual compression ratio is provided in an engine, and the actual compression ratio is changed according to an operation state.
Japanese Patent Laying-Open No. 2005-076579 JP 2005-146991 A “Internal Combustion Engine Fundamentals” by John B. Heywood

  In recent years, as a result of earnest research, the present inventors have increased the compression ratio so that the ignition timing determined from the knocking limit is after the compression top dead center (compression ratio = 13 or more), and the ignition timing retarded to prevent knocking. As the amount of retarded is reduced, the reduction in output due to the retard of the ignition timing is compensated by the thermal efficiency improvement, and it is possible to obtain as much fuel efficiency as a diesel engine without sacrificing the output. , By retarding the ignition timing within a predetermined period after compression top dead center, the effective compression ratio obtained at the intake valve closing timing defined by the valve lift of 1 mm at least in the high load operation region including the throttle fully open region in the low speed region Has invented a gasoline engine capable of maintaining 13 or more (Japanese Patent Application No. 2006-1009) 2).

  In practical use of such a high compression ratio engine, it is required to avoid knocking with an inexpensive configuration and improve emissions even in an operation region where reduction of the effective compression ratio is unavoidable. That is, when the geometric compression ratio is high, it is necessary to set the length of the exhaust system large in order to perform scavenging by overlapping the opening timings of the intake valve and the exhaust valve. However, as the length of the exhaust system increases, the time for the catalyst to reach the activation temperature increases, which may hinder emissions.

  On the other hand, providing a mechanism for changing the actual compression ratio as disclosed in Patent Documents 1 and 2 complicates the engine and increases the cost.

  The present invention has been made in order to solve the above-mentioned problems, and in the practical application of a high compression ratio engine, a spark ignition direct injection gasoline capable of avoiding knocking and improving emission with an inexpensive configuration. The challenge is to provide an engine.

In order to solve the above-described problems, the present invention provides a spark ignition gasoline engine having at least a spark plug and a fuel injection valve, an engine body having a plurality of cylinders and having a geometric compression ratio set to 14 or more. A catalyst unit that purifies the burned gas discharged from the engine body, and an intake valve and an exhaust valve that are respectively provided in an intake port and an exhaust port connected to each cylinder of the engine body and open and close the corresponding ports; A valve operating mechanism capable of changing the opening and closing timings of the intake valve and the exhaust valve, an exhaust pipe connected to the exhaust port and guiding burned gas from the exhaust port to the catalyst unit, and an operating state of the engine body. Based on detection of the operating state detecting means to detect and the operating state detecting means, at least the valve operating mechanism and fuel injection of the fuel injection valve Comprising: a timing, the ignition timing of the spark plug, and a control means for executing the adjustment control of an effective compression ratio by closing timing adjustment control of the intake valve, the fuel injection valve injects fuel around the spark plug The exhaust pipe is set to have a total path length of 500 mm or more from each exhaust port to the catalyst unit, and the control means is configured to take intake air when the engine is cold started. A phase difference is provided between the closing timing of the exhaust valve and the opening timing of the intake valve so that negative pressure is generated in the cylinder in the first half of the stroke, and at the intake valve closing timing defined by the valve lift of 1 mm. The valve mechanism is controlled so that the closing timing of the intake valve is in the vicinity of the intake bottom dead center so that the required effective compression ratio is maintained at 13 or more. Start timing and controls the fuel injection valve such that prior to the closing timing of the end timing of the fuel injection intake valve with becomes later the opening timing of a and the intake valve during the intake stroke, and the ignition timing is compression top dead in Ru spark-ignition gasoline engine der, characterized in that controls the spark plug so as to be retarded within a predetermined period after the point.

In this aspect, since the exhaust pipe of low exhaust pressure specification with the total path length set to 500 mm or more is connected to each exhaust port of the engine body, the exhaust gas is discharged when the burned gas is discharged from the plurality of cylinders. Interference is less likely to occur and knocking can be prevented from occurring. By the way, when the exhaust pipe is set to a low exhaust pressure specification (500 mm or more), it is assumed that the burned gas dissipates heat before reaching the catalyst unit, so that it is difficult for the catalyst unit to achieve the light-off temperature. However, in this aspect, when the engine is cold started, a phase difference is generated between the closing timing of the exhaust valve and the opening timing of the intake valve so that negative pressure is generated in the cylinder in the first half of the intake stroke. Since it is provided, a large negative pressure is generated in the cylinder in the first half of the intake stroke, and when the intake valve is opened, fresh air is sucked into the cylinder all at once. Therefore, an adiabatic compression effect is produced in the cylinder, which contributes to an increase in the cylinder temperature. In addition, since fuel injection is started after the intake valve is opened and fuel injection is terminated before the intake valve is closed, the turbulence in the cylinder is strengthened by the intake air flowing in force, and the injection Mixing of fuel and air is improved and vaporization of the injected fuel is promoted. In particular, since the geometric compression ratio of the engine body is set to 14 or more, a large negative pressure can be generated, and the adiabatic compression effect when the intake valve is opened is extremely high. In addition, in the engine body having a geometric compression ratio of 14 or more, the intake valve closing timing is set to the intake bottom dead center so that the effective compression ratio required at the intake valve closing timing defined by the valve lift of 1 mm is maintained at 13 or more. Since the pressure is controlled in the vicinity, the pressure in the cylinder is also increased, and the combustibility can be further enhanced. As a result, not only the output of the engine is improved, but also the ignition timing can be retarded in addition to the increase in the in-cylinder temperature due to the adiabatic compression action, and the exhaust gas temperature can be increased. For this reason, although the exhaust pipe of low exhaust pressure specification is adopted, the temperature rise of the entire exhaust system is promoted, and the time for the catalyst unit to reach the light-off temperature can be shortened. In addition, since the ignition retardability is improved due to the high temperature and high pressure in the cylinder, the ignition timing can be retarded more, the exhaust temperature can be increased, and the temperature increase rate of the catalyst unit can be increased. In addition, since the effective compression ratio is determined by the closing timing adjustment control of the intake valve, it is not necessary to use a complicated mechanism for changing the geometric compression ratio.

  In a preferred aspect, the control means controls the fuel injection valve so as to end the fuel injection in the first half of the intake valve opening period. In this mode, since fuel is injected when the negative pressure is further increased, mixing of the injected fuel and fresh air is further improved by the intake air flow, the combustion speed is improved, and the ignition retardability is improved. Will improve.

  In a preferred aspect, the control means injects the fuel into an intake stroke and a compression stroke so as to prevent an excess fuel atmosphere around the spark plug while maintaining a predetermined total air-fuel ratio. In this aspect, the penetration (penetration force) of the fuel spray is suppressed by the divided injection, the increase of HC and CO during cold due to the adhesion of the fuel spray to the wall surface is suppressed, and the relative fuel is injected by the fuel injected in the intake stroke. Since the amount of fuel injected in the compression stroke is reduced, the surrounding area of the spark plug becomes excessively excessive, or the fuel droplets are less likely to adhere to the spark plug, thereby preventing the spark plug from being swung.

  In the aspect of performing the split injection, the control means changes the total air-fuel ratio from rich to the stoichiometric air-fuel ratio over time after the start of the cold start operation, and even after the total air-fuel ratio becomes the stoichiometric air-fuel ratio. Continue compression injection. In this aspect, when realizing stratified combustion in an excess fuel atmosphere to generate hydrogen, the total fuel injection amount is suppressed just before the end of the cold start operation, so that the fuel injection amount is saved sufficiently and sufficiently. It becomes possible.

  In the aspect in which the split injection is performed, the control means sets the fuel injection ratio in the compression stroke to be larger than the fuel injection ratio in the intake stroke at the start of the cold start operation, and over time in the compression stroke. The fuel injection ratio is gradually decreased to increase the fuel injection ratio in the intake stroke. In this aspect, when realizing stratified combustion in an excess fuel atmosphere in order to generate hydrogen, the total fuel injection amount is suppressed, so that the fuel injection amount can be saved sufficiently.

  In a preferred embodiment, the control means is provided with a variable valve timing mechanism capable of changing at least the closing timing of the exhaust valve by the control of the control means, and the control means is in the high speed and high load operation region in the low speed and high load operation region of the engine The valve opening timing of the intake valve and the exhaust valve is controlled to be longer than the overlap. In this aspect, scavenging of the burned gas is promoted in the low speed and high load operation region, and the knocking resistance is improved. Further, by increasing the knocking resistance, it becomes possible to maintain a higher effective compression ratio.

In a preferred aspect, the exhaust pipe has an independent exhaust pipe connected to an exhaust port for each cylinder and a downstream end of an independent exhaust pipe provided in a cylinder whose combustion order is not adjacent to each other on the upstream side of the catalyst unit. those having a set portion communicating. In this aspect, since the volume of the exhaust pipe line increases from the upstream end of the exhaust pipe, the exhaust interference can be more reliably suppressed, and the burned gas HC and CO react with oxygen during the cold start operation. Therefore, a necessary volume can be secured.

  In a preferred aspect, the exhaust pipe has a plurality of the collecting portions, and each collecting portion is gathered upstream of the catalyst unit.

  As described above, the present invention has a remarkable effect that, when a high compression ratio engine is put into practical use, knocking can be avoided and emission can be improved with an inexpensive configuration.

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

  FIG. 2 is a configuration diagram illustrating a schematic configuration of a four-cycle spark ignition gasoline engine according to an embodiment of the present invention, and FIG. 3 is a schematic cross-sectional view illustrating a structure of one cylinder of the engine body 20 according to FIG. It is.

  2 and 3, the illustrated spark ignition type gasoline engine includes a four-cylinder in-line engine body 20 and an engine control unit 100 for controlling the engine body 20.

  The engine body 20 integrally includes a cylinder block 22 that rotatably supports the crankshaft 21 and a cylinder head 23 that is disposed above the cylinder block 22. Are provided with a plurality of cylinders 24.

  Each cylinder 24 is provided with a piston 26 connected to the crankshaft 21 via a connecting rod 25 and a combustion chamber 27 formed in the cylinder 24 by the piston 26. In the present embodiment, the geometric compression ratio of each cylinder 24 is set to 14. Further, if each cylinder 24 is defined as the first cylinder to the fourth cylinder in order from the left in FIG. 2, the combustion order is the order of the first cylinder, the third cylinder, the fourth cylinder, and the second cylinder. .

  Referring to FIG. 3, in the engine body 20 according to the present embodiment, the cylinder bore center Z of the cylinder 24 is the crankshaft 21 when viewed from the side where the rotation direction of the crankshaft 21 is clockwise (that is, the state of FIG. 3). Is offset to the right from the rotation center O. This offset amount is set to, for example, 1 mm to 2 mm when the bore diameter of the cylinder 24 is 70 mm.

  On the lower surface of the cylinder head 23, a ceiling portion of the combustion chamber 27 is formed for each cylinder 24, and this ceiling portion is a so-called pent roof type having two inclined surfaces extending from the central portion to the lower end of the cylinder head 23.

  A multi-hole injector 32 of a fuel injection system that receives a fuel injection pulse from the engine control unit 100 and injects fuel corresponding to the pulse width into the combustion chamber 27 is provided at the side of the combustion chamber 27. .

  FIG. 4 is an explanatory view for explaining the fuel injection direction of the multi-hole injector 32.

  Referring to FIG. 4, the multi-hole injector 32 has six injection holes 32 </ b> A to 32 </ b> F. Of these, the injection holes 32A and 32B have directivity for injecting fuel to both sides of the electrode of the spark plug 34, and the injection hole 32C has directivity for injecting fuel directly below the electrode. In the present embodiment, these three nozzle holes 32A to 32C constitute electrode-side nozzle holes.

  On the other hand, the injection port 32D has a directivity for injecting fuel further below the electrode than the injection port 32C, and the remaining injection ports 32E and 32F have a directivity for injecting fuel to both sides of the injection port 32D. In the present embodiment, these three nozzle holes 32D to 32F constitute a piston-side nozzle hole. In addition, in FIG. 4, the code | symbol of the spray injected from each nozzle hole 32A-32F is set to Fa-Ff.

  Each cylinder 24 is provided with a spark plug 34 that is fixed to the cylinder head 23 and emits a spark in the combustion chamber 27. The spark plug 34 is disposed on the cylinder bore center Z along the cylinder diameter parallel to the ridge line portion of the piston 26. An ignition circuit 35 capable of controlling the ignition timing by electronic control is connected to the ignition plug 34. By controlling the ignition circuit 35 by the engine control unit 100, the ignition plug 34 is controlled at a predetermined timing. The ignition is controlled.

  Next, each intake valve 30 is configured to be driven by a valve operating mechanism 40. The valve operating mechanism 40 has a variable camshaft timing mechanism (VCT) 36 that can change the opening / closing timing of the intake valve 30 steplessly, and a VVE (Variable) that can change the lift amount (opening amount) of the intake valve 30 steplessly. Valve Event). In addition, since VCT36 and VVE itself can employ | adopt a well-known mechanism, detailed description is abbreviate | omitted here. Further, the valve operating mechanism 40 as described above is also provided in the exhaust valve 31, and the closed timing of the exhaust valve 31 is advanced from the exhaust top dead center so that the burned gas in the cylinder remains in the intake stroke transition. It is possible to make it an internal EGR. In the illustrated example, only the VCT 36 is provided.

  Next, the branch intake pipe 133 of the intake manifold 132 is connected to the intake port 28 of the engine body 20. The branch intake pipe 133 is provided for each cylinder 24, and each branch intake pipe 133 is connected to the intake manifold 132 in a state where an equal-length intake path is formed. An intake passage 134 for introducing fresh air into the intake manifold 132 is connected to the upstream side of the intake manifold 132. A throttle valve 135 is provided in the intake passage 134. An air flow meter 136 is provided upstream of the throttle valve 135 in the intake passage 134.

  FIG. 5 is a schematic view taken along the line B-B in FIG. 2.

  With reference to FIGS. 2 and 5, the engine body 20 is provided with an exhaust pipe 140. The exhaust pipe 140 is connected to the exhaust port 29 and has a bifurcated independent exhaust pipe 141 formed in pairs in each cylinder 24. As described above, in the present embodiment, the combustion order of each cylinder 24 is the order of the first cylinder, the third cylinder, the fourth cylinder, and the second cylinder. Accordingly, the exhaust pipe 140 according to the present embodiment communicates the downstream ends of the independent exhaust pipes 141 provided in the first cylinder and the fourth cylinder to form a first collecting portion 142, and the second cylinder and the third cylinder. The exhaust manifold 144 is formed as a second collecting portion 143 by communicating with the downstream end of the independent exhaust pipe 141 provided in the first and the second. That is, each of the collecting portions 142 and 143 communicates with the independent exhaust pipe 141 connected to the cylinders whose combustion order is not adjacent (not continuous). As a result, when the burned gas is discharged from any of the cylinders 24, the exhaust gas is not discharged from the plurality of cylinders 24 at the same time, and the exhaust gas interference phenomenon (exhaust gas is simultaneously discharged into the same exhaust path). By discharging, it is possible to avoid the phenomenon that the flow of exhaust gas is hindered in the exhaust manifold 144).

Each independent exhaust pipe 141 is set to a relatively long path length L1 (for example, 300 mm). As a result, even when scavenging of the combustion chamber 27 is executed, the exhaust gas is discharged quickly, and even when the effective compression ratio is set high, knocking can be effectively avoided. It has become. In the present embodiment, the entire path length (the length from the exhaust port 29 to the downstream ends of the collecting portions 142 and 143 ) L is set to about 600 mm, which is longer than 500 mm, for example. The exhaust system including each independent exhaust pipe 141 and the exhaust manifold 144 extends from the lower part of the dash panel 151 to the lower part of the floor panel 152 that is continuous with the dash panel 151.

  Referring to FIG. 5, catalyst unit 146 is connected to exhaust pipe 145 connected to the downstream side of exhaust manifold 144, and is fixed to the bottom surface of floor panel 152 of the vehicle. The catalyst unit 146 includes a three-way catalyst 146a and a Pt / Rh catalyst 146b that is disposed downstream of the three-way catalyst 146a and has a NOx storage function as a NOx catalyst. In FIG. 5, 153 is a hood panel that covers the engine body 20, and 154 is a transmission unit attached to the engine body 20.

  Referring to FIG. 3, in order to detect the operating state of engine body 20, cylinder block 22 detects a pair of crank angle sensors SW1 and SW2 that detect the rotational speed of crankshaft 21 and the temperature of cooling water. An engine water temperature sensor SW3 is provided (see FIG. 3). Further, the exhaust passage 153 is provided with an air-fuel ratio sensor SW4 for controlling the air-fuel ratio.

  The engine body 20 is provided with an engine control unit 100 as control means. The engine control unit 100 is connected with an air flow meter 136, crank angle sensors SW1, SW2, an engine water temperature sensor SW3, an air-fuel ratio sensor SW4, and an accelerator opening sensor SW5 for detecting an engine load as input elements. . The air flow meter 136 and the sensors SW1 to SW5 are all specific examples of the operation state detection sensor in the present embodiment. On the other hand, the engine control unit 100 is connected with a fuel injection system of the multi-hole injector 32, a spark plug 34, a valve mechanism 40, and the like as control elements.

  The engine control unit 100 includes a microprocessor having a CPU, a memory, an interface, a counter / timer group, and a bus connecting these units. The engine control unit 100 functionally configures an operation state determination unit that determines an operation state based on a program and data stored in a memory.

  The memory of the engine control unit 100 stores various control maps, which will be described in detail later. Based on these storage maps, the engine body 20 is preferably operated according to the operating state. .

  FIG. 6 is a graph showing the relationship between the engine speed N that is the basis of the control map and the required torque in the embodiment of FIG.

Referring to FIG. 6, the illustrated embodiment includes an idling operation region A, a low / medium / low / medium load operation region B, a low / high load operation region C, and a high / high load operation region D. Among these, the idling operation region A is an idling operation region up to a predetermined rotation speed N1 (for example, 500 rpm). Further, the low, medium, low, and medium load operation region B is divided into three regions of N1, N2 (for example, 2000 rpm), and N3 (for example, 4000 rpm), and the low, medium, and medium speed region ( In the range up to N2) and is set in the range of the low and medium load operation region excluding the idling operation region. Further, the low-speed and high-load operation region C is in the low-speed region when the region of the engine rotation speed N is divided into three stages at the rotation speeds N2 and N3, and the boundary portion of the low-medium-speed and low-medium-load operation region B To the throttle fully open range C _WOT .

  FIG. 7 is a graph showing an example of the ignition timing that is the basis of the control map in the embodiment of FIG.

Referring to FIG. 7, for example, when the geometric compression ratio is 11, the ignition timing during normal operation is a considerable amount A _Ig (for example, the engine speed is higher than the compression top dead center, as indicated by IG _a ). 1500 rpm, crank angle CA = 6 ° to 8 °). In contrast, if the geometric compression ratio is 14, if the same knock characteristics as compression ratio 11 as shown by IG _v, but is was not ignited immediately before the compression top dead center, in this embodiment , as shown by IG _b, is set to be spark ignited at ignition timing is further retarded than the compression top dead center. Thereby, in this embodiment, in the retard operation region (low speed medium and high load operation region including the throttle full open region _CWOT ) C, the state where the torque is not reduced and the high compression ratio (ε _r ≦ 14) is still maintained. (See FIG. 7).

When the ignition timing is retarded after compression top dead center, the retard amount R _Ig is experimentally accumulated in consideration of factors that determine knocking, such as in-cylinder temperature and in-cylinder pressure, and is determined by a control map. However, in the present embodiment, for example, the retard amount R _Ig from the compression top dead center is set to a stroke range of 10% or less after the top dead center has elapsed (crank angle CA = around 35 ° after compression top dead center). . By retarding the ignition timing after the compression top dead center, knocking is suppressed and operation at a high compression ratio is possible, but combustion is performed as much as the ignition timing is retarded from the compression top dead center. It is disadvantageous in terms of time. Therefore, in the present embodiment, in order to combust the fuel that has been able to suppress knocking and has shifted to the expansion stroke at an early stage, the retard amount R _{Ig is set} to a stroke range in which the piston 26 is 10% or less after the top dead center has elapsed. It is.

  FIG. 8 is a graph showing the air-fuel ratio and the concentration of the exhaust gas component for explaining the exhaust gas components, and FIG. 9 shows the air-fuel ratio and the count time Tc that are the basis of the control map in the embodiment of FIG. It is a graph which shows the relationship.

Referring to FIG. 8, when the air-fuel ratio is rich, a lot of HC and CO are generated as harmful components contained in the exhaust gas. At the same time, it is known that a large amount of hydrogen (H ₂ ) is generated.

  Therefore, in the present embodiment, as shown in FIG. 9, a predetermined count time Tc is set in the cold start operation described later, and the in-cylinder air-fuel ratio is set rich when the cold start is started. (For example, A / F≈10), and after starting the cold start, after the count time Tc has elapsed, control is performed so that the fuel injection amount is gradually decreased so that the total air-fuel ratio in the cylinder becomes the stoichiometric air-fuel ratio. A map M2 (see FIG. 11) is created.

  Next, a control example of the engine of FIG. 2 will be described.

  10 to 12 are flowcharts showing an example of control of the engine of FIG.

Referring to FIG. 10, engine control unit 100 reads each data (step S20) and determines the operating state of the engine (step S21). Next, the engine control unit 100 indexes the effective compression ratio ε _r to be set based on the predetermined control map M1 (step S22). Control map M1 for determining the effective compression ratio epsilon _r is the rotational speed of the engine, the required load, vehicle speed, based on the in-cylinder temperature and the like, are set such that the effective compression ratio epsilon _r is determined.

Here, if the effective compression ratio is ε _r ,

(2) where
ε : Geometric compression ratio v _s : Stroke volume (m ³ )
v _c : Clearance volume (m ³ )
θ : Crank angle R at the closing timing of the intake valve 30 when the valve lift is 1 mm : Continuous ratio (connector rod length / crank radius)
It is.

By using the equation (2), the relationship between the effective compression ratio ε _r and the valve opening angle is converted into data based on the valve opening angle of the intake valve 30 when the valve lift is 1 mm, and the control map M1 By doing so, it becomes possible to precisely control the effective compression ratio ε _r .

Next, the engine control unit 100 sets the closing timing of the intake valve 30 corresponding to the indexed effective compression ratio ε _r (step S23), and controls the valve mechanism 40 based on this setting. Here, in the present embodiment, during the cold start of the engine, between the closing timing of the exhaust valve 31 and the opening timing of the intake valve 30 so that negative pressure is generated in the cylinder in the first half of the intake stroke. A valve is provided so that the closing timing of the intake valve 30 is in the vicinity of the intake bottom dead center so that the effective compression ratio ε _r obtained at the intake valve closing timing defined by the valve lift of 1 mm is maintained at 13 or more. The mechanism 40 is configured to be controlled. By executing such control, it is possible to realize adiabatic compression start (operation control for starting the engine by the adiabatic compression effect in the intake stroke at the cold start).

  The engine control unit 100 determines whether or not the cold operation condition is satisfied based on the operation state determined in step S21 (step S24). Here, the cold operation condition refers to a case where the operation region is in the idling operation region A shown in FIG. 6 and the in-cylinder temperature state is low (for example, the engine coolant temperature is 40 ° C.) or less.

  When it is determined that the engine operating condition satisfies the cold operating condition, the engine control unit 100 executes the routine shown in FIG. 11 from step S25 to step S33 to promote the activation of the catalyst unit 146. Then, cold start control is executed.

  Referring to FIG. 11, engine control unit 100 starts counting count time Tc described in FIG. 9 and cycle number Cy of engine body 20 at the same time (step S25). Next, the air-fuel ratio is indexed from the operating state and Tc based on the control map M2 described in FIG. 9 (step S26). Next, the engine control unit 100 indexes the fuel injection amount with the air-fuel ratio indexed from the predetermined control map M3 as the target air-fuel ratio (step S27). The control map M3 is set so that the fuel injection amount can be determined by the intake air flow rate, the engine rotation speed, and the like. As described above, in the cold start control, the total air-fuel ratio in the cylinder at the start of operation is set to be relatively rich (A / F = 10 in the example of FIG. 9). As a result, since the multi-hole type injector 32 having the electrode side injection holes 32A to 32C is employed as the fuel injection valve, the stratified combustion around the spark plug is generated in the combustion chamber 27 at the time of ignition. .

  Next, the engine control unit 100 determines the fuel split ratio and fuel injection timing based on the indexed fuel injection amount (step S28). In the present embodiment, the fuel injection timing is divided into an intake stroke and a compression stroke during the cold start (see F1 and F2 in FIG. 13A). As a result, stratified combustion in the cylinder can be realized with certainty, and fuel injection is performed even in the intake stroke, so that the atmosphere around the spark plug 34 becomes excessively excessive or the injected fuel is completely discharged. It becomes possible to suppress that.

Next (or in parallel with the setting of the fuel injection timing or the like), the engine control unit 100 indexes the ignition timing based on the control map M4 (step S29). Control map M4 has been set by an experiment or the like based on the graph of FIG. 7, on the basis of the effective compression ratio epsilon _r, is configured to be able to determine the ignition timing. When the cold operation condition is satisfied, the operation region is the idling operation region A in FIG. 6, and therefore the effective compression ratio ε _r set in step S23 is maintained at 13 or more. Accordingly, in step S29, the spark plug 34 is operated at a relatively early timing after the compression top dead center has elapsed.

  Next, the engine control unit 100 operates the fuel injection system at the set timing to execute fuel injection control (step S30), and operates the spark plug 34 after the compression top dead center has elapsed to execute ignition control. (Step S31).

  As described above, in the cold start control, stratified combustion occurs in the combustion chamber 27 at the time of ignition. For this reason, the exhaust gas discharged into the exhaust system after combustion contains a relatively large amount of hydrogen together with HC and CO. This hydrogen reacts with surplus oxygen in the path from the independent exhaust pipe 141 to the catalyst unit 146. As a result, the generated oxidation heat oxidizes HC and CO to purify the exhaust gas, and the oxidation heat raises the temperature of the low exhaust pressure specification exhaust system set to a long path length. . As a result, the exhaust gas containing reaction heat is introduced into the catalyst unit 146 to promote the catalyst unit 146 reaching the light-off temperature. As a result, HC and CO remaining in the exhaust gas are easily purified by a catalytic reaction by the three-way catalyst 146a.

  Thereafter, the engine control unit 100 determines whether or not a cold operation end condition is satisfied (step S32). Here, the cold operation end condition is a condition for ending the cold start control. Specifically, the number of cycles counted after executing the cold start control is a predetermined value (for example, The determination is made based on whether or not 5, 6) is reached. In addition, or in parallel with this, a temperature sensor is provided in the catalyst unit 146, and it is determined that the cold operation end condition is satisfied depending on whether or not the catalyst unit 146 has reached the light-off temperature. May be.

  If it is determined in step S32 that the cold operation end condition is satisfied, the engine control unit 100 resets the count time Tc and the cycle number Cy (step S33), returns to step S20, and performs the subsequent steps. repeat.

  If it is determined in step S32 that the cold operation end condition is not satisfied, the engine control unit 100 continues counting the count time Tc (step S34), increments the cycle number Cy (step S35), and Control goes to step S26. When the control shifts to step S26 after step S35, the air-fuel ratio indexed from the control map M2 by the engine control unit 100 is leaner than the previous set value because Tc has elapsed. After a predetermined time Tn has elapsed, the stoichiometric air-fuel ratio is set.

  Next, when it is determined in step S24 of FIG. 10 that the cold operation condition is not satisfied (NO in step S24), the engine control unit 100 executes the flow shown in FIG.

  Referring to FIG. 12, in this flow, the air-fuel ratio is indexed from the operating state based on control map M2 described in FIG. 9 (step S40). In order to realize this control, an area for storing a cold start control execution flag is provided in the memory of the engine control unit 100, and when the execution flag is ON, an air-fuel ratio that takes the count time Tc into account as a parameter. When the execution flag is OFF, the control time as described above is performed using the same control map M2 by setting the air-fuel ratio to be indexed by removing the count time Tc from the parameter. It is possible to realize.

  Next, the engine control unit 100 indexes the fuel injection amount with the air-fuel ratio indexed from the control map M3 as the target air-fuel ratio (step S41).

  In normal operation control, the total air-fuel ratio in the cylinder is set to lean that is relatively close to the stoichiometric air-fuel ratio.

  Next, the engine control unit 100 determines the split ratio of fuel to be injected and the fuel injection timing based on the indexed fuel injection amount (step S42).

Next (or in parallel with the setting of the fuel injection timing or the like), the engine control unit 100 executes control for determining the ignition timing. In this process, the engine control unit 100 determines whether or not the effective compression ratio ε _r set in step S23 is 13 or more (step S43), and when the effective compression ratio ε _r is 13 or more. The ignition timing is retarded at once to the maximum allowable value after the compression top dead center (step S44), and when the effective compression ratio ε _r is less than 13, the predetermined timing before the compression top dead center based on the control map M4. The ignition timing is set to (step S45).

  Next, the engine control unit 100 operates the fuel injection system at the set timing to execute fuel injection control (step S46), and operates the spark plug 34 before compression top dead center to execute ignition control. (Step S47).

  Thereafter, the engine control unit 100 returns to step S20 and repeats the subsequent steps.

  FIG. 13 is a timing chart showing a control example when the flowcharts of FIGS. 10 to 12 are executed. (A) is a cold start (adiabatic compression start), (B) is a warm low speed operation, (C) is during medium and high speed high load operation, and (D) is during low speed and high load operation.

Next, with reference to FIG. 13, the valve opening timing, the effective compression ratio ε _r , and the ignition timing of the intake valve 30 during each operation will be described in detail. The advance / retard operation of the intake valve 30 is realized by the VCT 36 of the intake valve 30, and the lift amount operation is realized by the VVT of the valve mechanism 40.

Referring to FIG. 13 (A), when the operation region is the idling operation region A in FIG. 6 and the cold start condition is satisfied, the opening timing of the intake valve 30 is The valve closing timing is set in the vicinity of the intake bottom dead center. Further, the lift amount of the intake valve 30 is set to be relatively small. Furthermore, the fuel injection during the intake stroke (F1) is initiated after the valve-opening timing of the intake valve 30, Ru is terminated before the closing timing of the intake valve 30. As a result, the effective compression ratio ε _r is generally maintained at a value (13 or more) close to the geometric compression ratio (14 in the present embodiment), and adiabatic expansion occurs in the cylinder in the first half of the intake stroke. become. Therefore, when the intake valve 30 is opened, the large negative pressure generated in the cylinder causes the fresh air or fuel (F1) injected in the intake stroke to be sucked into the cylinder. Thereby, the adiabatic compression effect is generated in the cylinder and contributes to an increase in the in-cylinder temperature. As a result, fuel atomization is promoted. In particular, since the geometric compression ratio of the engine body 20 is set to 14 or more, a large negative pressure can be generated, and the adiabatic compression effect when the intake valve 30 is opened becomes extremely high. Further, as the effective compression ratio ε _r is maintained at a value substantially close to the geometric compression ratio, in this operating state, the ignition timing is retarded after the compression top dead center. Thereby, knocking is reliably prevented. In this embodiment, since the ignition retardability is improved by the high temperature and high pressure in the cylinder, the ignition timing is retarded more, the exhaust temperature is increased, and the temperature increase rate of the catalyst unit 146 can be increased. it can. In addition, since stratified combustion is realized by the fuel (F2) injected in the compression stroke, the amount of generated hydrogen is increased, and the reaction between hydrogen and harmful components in the exhaust pipe 140 is promoted in the cold state. The reaction heat can increase the temperature of the entire exhaust system and contribute to the activation of the catalyst unit 146. Further, by executing fuel injection in the compression stroke, in combination with the fact that the geometric compression ratio is set high, the sprays Fa ~ injected around the spark plug 34 from the electrode side injection holes 32A to 32C. The penetration of Fc is shortened, and the atmosphere around the spark plug 34 becomes a fuel excess atmosphere more effectively. Further, as the penetration becomes shorter due to the increase in the in-cylinder pressure in the compression stroke, the time for staying around the spark plug 34 also becomes longer, so that the ratio of the unburned gas discharged becomes higher. Therefore, in combination with the generation of hydrogen, the combustion temperature in the exhaust system can be increased. On the other hand, the amount of fuel injected in the compression stroke is relatively reduced by the fuel injected in the intake stroke, so that the atmosphere around the spark plug 34 becomes excessively excessive, or the fuel droplets on the spark plug 34 Becomes difficult to adhere.

  Referring to FIG. 13B, when the operation region is the low, medium speed, and low load operation region B of FIG. 6 and is in the warm operation, the intake valve 30 has an opening / closing timing in the latter half of the exhaust stroke. The closing timing is set so as to advance in the latter half of the intake stroke. In the setting control in step S26, the lift amount is set to a large value in accordance with the required intake air amount (which increases as the engine load and the rotational speed increase), the closing timing is advanced, and the pumping loss is reduced. I try to figure it out. As the intake valve 30 is closed early, the ignition timing is also advanced before the compression top dead center.

Referring to FIG. 13C, during low-speed and high-load operation (low-speed and high-load operation region C in FIG. 6), the intake valve 30 is advanced and lifted as compared to the operation region in FIG. By increasing the amount, the closing timing is set immediately after the compression bottom dead center has elapsed. Further, the exhaust valve 31 is retarded as compared with the operation region of FIG. 13B, and the closing timing is set in the first half of the intake stroke. Thereby, the valve opening timings of the intake valve 30 and the exhaust valve 31 overlap from the latter half of the exhaust stroke to the first half of the intake stroke, and scavenging of the combustion chamber 27 is promoted. Here, in the present embodiment, the independent exhaust pipe 141 having a relatively long path length L1 is adopted to constitute a low exhaust pressure specification exhaust system having a considerably long path overall length L, and the first cylinder and the fourth cylinder. The downstream end of the independent exhaust pipe 141 provided in the cylinder communicates with the first collecting portion 142, and the downstream end of the independent exhaust pipe 141 provided in the second and third cylinders is the second collecting portion 143. An exhaust manifold 144 that communicates with the exhaust manifold 144 is employed. Therefore, when the burned gas is discharged from any cylinder 24, particularly when scavenging is performed as shown in FIG. 13C, the exhaust gas is not discharged from the plurality of cylinders 24 at the same time. Gas interference phenomenon can be avoided. Further, in the operation region of FIG. 13C, the effective compression ratio ε _r is maintained at a value substantially close to the geometric compression ratio, so that the ignition timing is retarded after the compression top dead center. In this embodiment, by retarding the ignition timing after the compression top dead center, knocking is avoided and high output is maintained while maintaining a high compression ratio even in the low speed and high load operation region including the throttle fully open region _CWOT. It can be obtained.

  Referring to FIG. 13 (D), in the high speed and high load operation region D of FIG. 6, the intake valve 30 is advanced and the lift amount is further increased compared to the operation region of FIG. 13 (B). The closing timing is set immediately after the compression bottom dead center has elapsed. On the other hand, the opening / closing timing of the exhaust valve 31 is advanced and the overlap of the valve opening timing is smaller than that in the operation region of FIG. Thereby, a high output is ensured. Note that the crank angle timing of the injection pulse becomes longer as the fuel injection timing increases as the engine speed increases.

As described above, in the present embodiment, since the engine body 20 has the exhaust pipe 140 set to a predetermined low exhaust pressure specification, when the burned gas is discharged from the plurality of cylinders 24, the exhaust gas is exhausted. Interference is less likely to occur and knocking can be prevented from occurring. Here, the low exhaust pressure specification is a specification having a longer path overall length than an exhaust pipe of a normal gasoline engine. For example, in the case of an in-line four-cylinder 24 engine, the exhaust port 29 to the catalyst unit 146 The total path length L is set to 500 mm or more (for example, about 600 mm) . By the way, when the exhaust pipe 140 is set to the low exhaust pressure specification, it is assumed that the burned gas radiates heat before reaching the catalyst unit 146, so that it is difficult for the catalyst unit 146 to achieve the light-off temperature. However, in this embodiment, when the engine is cold-started, the exhaust valve is closed between the closing timing of the exhaust valve and the opening timing of the intake valve 30 so that negative pressure is generated in the cylinder in the first half of the intake stroke. Since the phase difference is provided, a large negative pressure is generated in the cylinder in the first half of the intake stroke, and when the intake valve 30 is opened, fresh air is sucked into the cylinder all at once. Therefore, an adiabatic compression effect is produced in the cylinder, which contributes to an increase in the cylinder temperature. Also, since the fuel injection after the opening of the intake valves 30 (F1) is started, it is enhanced turbulence intensity in the cylinder by vigorously flows to the intake air flow, thereby improving the mixing of the injected fuel and air The vaporization and atomization of the injected fuel is promoted. In particular, since the geometric compression ratio of the engine body 20 is set to 14 or more, a large negative pressure can be generated, and the adiabatic compression effect when the intake valve 30 is opened becomes extremely high. In addition, in the engine body 20 having a geometric compression ratio of 14 or more, the closing timing of the intake valve 30 is set so that the effective compression ratio required at the closing timing of the intake valve 30 defined by the valve lift of 1 mm is maintained at 13 or more. Since the pressure is controlled in the vicinity of the bottom dead center, the pressure in the cylinder also increases, and the combustibility can be further improved. As a result, not only the output of the engine is improved, but also the ignition timing can be retarded in addition to the increase in the in-cylinder temperature due to the adiabatic compression action, and the exhaust gas temperature can be increased. For this reason, although the low exhaust pressure specification exhaust pipe is employed, the temperature rise of the entire exhaust system is promoted, and the time for the catalyst unit 146 to reach the light-off temperature can be shortened. In addition, since the ignition retardability is improved by the high temperature and high pressure in the cylinder, the ignition timing is further retarded, the exhaust temperature is increased, and the temperature increase rate of the catalyst unit 146 can be increased. In addition, since the effective compression ratio is determined by the closing timing adjustment control of the intake valve 30, it is not necessary to use a complicated mechanism for changing the geometric compression ratio.

  Further, in the present embodiment, the multi-hole injector 32 as a fuel injection valve is controlled so as to end the fuel injection within the valve opening period of the intake valve 30. For this reason, in this embodiment, the mixing of the injected fuel and fresh air becomes good due to the intake air flow, the combustion speed is improved, and the ignition retardability is improved.

  In the present embodiment, the multi-hole injector 32 is controlled so that fuel injection is terminated in the first half of the opening period of the intake valve 30. For this reason, in the present embodiment, fuel is injected when the negative pressure is further increased. Therefore, the mixing of the injected fuel and fresh air is further improved by the intake air flow, the combustion speed is improved, and the ignition is performed. The retardability is improved.

  Further, in this embodiment, a direct injection type multi-hole injector 32 having electrode side injection holes 32A to 32C for injecting fuel around the spark plug 34 is adopted, and the periphery of the spark plug 34 is maintained while maintaining a predetermined total air-fuel ratio. In order to avoid an excessive fuel atmosphere, the fuel is divided and injected into an intake stroke and a compression stroke. For this reason, in this embodiment, penetration (penetration force) of fuel spray is suppressed by split injection, and an increase in HC and CO during cold due to adhesion of fuel spray to the wall surface is suppressed and fuel injected in the intake stroke is used. Since the fuel injection amount injected in the compression stroke is relatively reduced, the surrounding area of the spark plug becomes excessively fuely, or the fuel droplets are less likely to adhere to the spark plug, thereby preventing the spark plug from being blown. it can.

  In this embodiment, after the cold start operation is started, the total air-fuel ratio is changed from rich to the stoichiometric air-fuel ratio over time, and compression injection is continued even after the total air-fuel ratio becomes the stoichiometric air-fuel ratio. is there. Therefore, in this embodiment, when realizing stratified combustion in an excess fuel atmosphere to generate hydrogen, the total fuel injection amount is suppressed just before the end of the cold start operation, so the fuel injection amount is necessary and sufficient. It becomes possible to save on.

  In the present embodiment, after the cold start control is started, the fuel injection amount is gradually decreased so that the total air-fuel ratio in the cylinder becomes the stoichiometric air-fuel ratio after a predetermined timing. For this reason, in this embodiment, when controlling the air-fuel ratio in the cylinder, immediately after the start of the cold operation, hydrogen is generated under rich conditions with a relatively low air-fuel ratio, and reacted with surplus oxygen to generate exhaust gas. While purifying and raising the temperature of the catalyst unit 146, as time elapses, the air-fuel ratio in the cylinder is made lean to suppress the generation of HC and CO. Purification and NOx occlusion can be achieved.

  Further, in the present embodiment, in the low speed and high load operation region C of the engine, the valve opening timing of the intake valve 30 and the exhaust valve 31 is controlled to overlap longer than in the high speed and high load operation region D. is there. For this reason, in the present embodiment, scavenging of burned gas is promoted in the low speed and high load operation region C, and the knocking resistance is improved. Further, by increasing the knocking resistance, it becomes possible to maintain a higher effective compression ratio.

  Further, in the present embodiment, the exhaust pipe 140 is connected to the exhaust port 29 for each cylinder 24, and is provided in the cylinder 24 where the combustion order is not adjacent to the independent exhaust pipe 140 that is set long for low exhaust pressure specifications. The independent exhaust pipe 140 has collecting portions 142 and 143 that communicate with downstream ends of the independent exhaust pipe 140 on the upstream side of the catalyst unit 146. For this reason, in the present embodiment, the volume of the exhaust pipe 140 from the upstream end of the exhaust pipe 140 is increased, so that it is possible to more reliably suppress the exhaust interference and to reduce the HC and CO of the burned gas during the cold start operation. The volume necessary for reacting with oxygen can be ensured.

  Therefore, according to this embodiment, when a high compression ratio engine is put to practical use, there is a remarkable effect that knocking can be avoided and emission can be improved with an inexpensive configuration.

  The above-described embodiments are merely preferred specific examples of the present invention, and the present invention is not limited to the above-described embodiments.

  FIG. 14 is a configuration diagram showing a schematic configuration of an engine according to another embodiment of the present invention.

  Referring to FIG. 14, for example, as an exhaust pipe 140 with a low exhaust pressure specification, the downstream ends of all the independent exhaust pipes 141 may be gathered and communicated with a single gathering part 142.

  FIG. 15 is a timing chart according to another embodiment of the present invention.

  Referring to FIGS. 15A and 15B, in embodying step S28 of FIG. 11, the fuel injection ratio in the intake stroke and the split stroke may be changed as the number of cycles Cy increases. . Specifically, as shown in FIG. 15A, at the start of the cold start operation, the fuel injection ratio in the compression stroke is set to be larger than the fuel injection ratio in the intake stroke, and the compression stroke over time is set. As shown in FIG. 15B, in the cycle number Cy (= n) just before the end of the cold start operation control, the fuel injection ratio in the intake stroke is changed in the compression stroke. This is higher than the fuel injection ratio.

  In the embodiment shown in FIG. 15 as well, when realizing stratified combustion in an excess fuel atmosphere to generate hydrogen, the total fuel injection amount is suppressed, so that the fuel injection amount can be saved sufficiently and sufficiently. Become.

  16 and 17 are timing charts according to still another embodiment of the present invention.

  As shown in FIG. 16, the fuel injection timing at the time of cold start is not limited to split injection, but may be only the intake stroke. That is, when the temperature setting at the time of cold start is relatively high, the in-cylinder temperature may be increased by the adiabatic compression effect by injecting fuel only in the intake stroke as shown in FIG. Further, when the control of FIG. 16 is adopted, it can be carried out even with a port injection type engine.

  In addition, as shown in FIG. 17, after the exhaust valve 31 is closed, the multi-hole injector is configured to start fuel injection after the intake top dead center and finish the fuel injection by the opening timing of the intake valve 30. 32 may be controlled. In that case, the fuel is injected when the negative pressure is generated in the cylinder, and as a result, the vaporization and atomization of the fuel by the negative pressure is promoted. As a result, the combustion speed is improved and the ignition retardance is reduced. improves.

Further, as a method of reducing the effective compression ratio ε _r , in the above-described embodiment, the valve operating mechanism 40 that can change the opening / closing timing of the intake valve 30 in a stepless manner is used. A valve mechanism with a so-called lost motion function that can switch the opening / closing timing of the intake valve 30 in two stages by selectively transmitting the cam to the intake valve 30 may be used.

  Further, the valve mechanism with the lost motion function as described above may be adopted as means for executing the internal EGR.

  It goes without saying that various modifications can be made within the scope of the claims of the present invention.

It is a graph which shows an example of the ignition retard at the time of high load driving | operation. It is a block diagram which shows schematic structure of the engine which concerns on one Embodiment of this invention. FIG. 3 is a schematic cross-sectional view taken along line AA showing the structure of one cylinder of the four-cycle spark ignition gasoline engine according to FIG. 2. It is explanatory drawing explaining the fuel injection direction of the multi-hole type injector. FIG. 3 is a schematic view taken along line BB in FIG. 2. It is a graph which shows the relationship between the engine speed used as the basis of a control map in the embodiment of FIG. 2, and a request torque. It is a graph which shows an example of the ignition timing used as the basis of a control map in embodiment of FIG. It is a graph which shows the air fuel ratio for demonstrating the component of exhaust gas, and the density | concentration of an exhaust gas component. 3 is a graph showing a relationship between an air-fuel ratio which is a basis of a control map and a count time in the embodiment of FIG. It is a flowchart which shows the example of control of the engine of FIG. It is a flowchart which shows the example of control of the engine of FIG. It is a flowchart which shows the example of control of the engine of FIG. FIG. 13 is a timing chart showing an example of control when the flowcharts of FIGS. 10 to 12 are executed, in which (A) is a cold start (adiabatic compression start), (B) is a warm low speed operation, and (C) is a (D) is during low-speed and high-load operation during medium and high-speed and high-load operation. It is a block diagram which shows schematic structure of the engine which concerns on another embodiment of this invention. 6 is a timing chart according to another embodiment of the present invention. 6 is a timing chart according to another embodiment of the present invention. 6 is a timing chart according to another embodiment of the present invention.

20 Engine Body 24 Cylinder 26 Piston 27 Combustion Chamber 28 Intake Port 29 Exhaust Port 30 Intake Valve 31 Exhaust Valve 32 Multi-Hole Injector 32A-32C Electrode Side Injection Port 34 Spark Plug 35 Ignition Circuit 36 VCT
40 Valve Control Mechanism 100 Engine Control Unit 140 Exhaust Pipe 141 Independent Exhaust Pipe 142 Collecting Portion 143 Collecting Portion 144 Exhaust Manifold 145 Exhaust Pipe 146 Catalyst Unit 151 Dash Panel 152 Floor Panel A Idling Operating Area B Low Medium Speed Low Medium Load Operating Area C Low speed and high load operation area C _WOT throttle fully open area D High speed and high load operation area L Total path length L1 Path length Tc Count time ε _r Effective compression ratio

Claims (9)

In a spark ignition gasoline engine having at least a spark plug and a fuel injection valve,
An engine body having a plurality of cylinders and having a geometric compression ratio of 14 or more;
A catalyst unit for purifying burned gas discharged from the engine body;
An intake valve and an exhaust valve that are respectively provided in an intake port and an exhaust port connected to each cylinder of the engine body, and open and close the corresponding ports;
A valve operating mechanism capable of changing the opening and closing timing of the intake valve and the exhaust valve;
An exhaust pipe that is connected to the exhaust port and guides burnt gas from the exhaust port to the catalyst unit;
Driving state detecting means for detecting the driving state of the engine body;
Based on the detection of the operating state detection means, at least the valve mechanism, the fuel injection timing of the fuel injection valve, the ignition timing of the ignition plug, and the effective compression ratio adjustment control by the intake valve closing timing adjustment control And a control means for executing
The fuel injection valve is a direct injection type having a port for injecting fuel around the spark plug,
The exhaust pipe has a total path length from each exhaust port to the catalyst unit set to 500 mm or more,
The control means provides a phase difference between the closing timing of the exhaust valve and the opening timing of the intake valve so that a negative pressure is generated in the cylinder in the first half of the intake stroke when the engine is cold started. And the valve mechanism is controlled so that the closing timing of the intake valve is close to the intake bottom dead center so that the effective compression ratio obtained at the closing timing of the intake valve defined by the valve lift of 1 mm is maintained at 13 or more. The fuel injection valve is controlled so that the fuel injection start timing is during the intake stroke and after the intake valve opening timing, and the fuel injection end timing is before the intake valve closing timing , and the ignition timing A spark ignition gasoline engine characterized in that the spark plug is controlled so as to be retarded within a predetermined period after compression top dead center. The spark ignition gasoline engine according to claim 1,
The spark ignition gasoline engine according to claim 1, wherein the control means controls the fuel injection valve so as to end the fuel injection in the first half of the intake valve opening period . The spark ignition gasoline engine according to claim 1 ,
The control means is a spark ignition type characterized in that fuel is divided and injected into an intake stroke and a compression stroke so as not to create an excessive fuel atmosphere around the spark plug while maintaining a predetermined total air-fuel ratio. gasoline engine. In the spark ignition gasoline engine according to claim 3 ,
The control means changes the total air-fuel ratio from rich to the stoichiometric air-fuel ratio over time after the start of the cold start operation, and continues the compression injection even after the total air-fuel ratio becomes the stoichiometric air-fuel ratio. This is a spark ignition gasoline engine. In the spark ignition gasoline engine according to claim 3 ,
At the start of the cold start operation, the control means sets the fuel injection ratio in the compression stroke to be larger than the fuel injection ratio in the intake stroke, and gradually reduces the fuel injection ratio in the compression stroke over time. A spark ignition gasoline engine characterized by increasing the fuel injection ratio in the stroke . In the spark ignition gasoline engine according to claim 3 ,
The spark ignition type characterized in that the control means gradually decreases the fuel injection amount so that the total air-fuel ratio in the cylinder becomes the stoichiometric air-fuel ratio after a predetermined timing after starting the cold start control. gasoline engine. The spark ignition type gasoline engine according to any one of claims 1 to 6 ,
A variable valve timing mechanism capable of changing at least the closing timing of the exhaust valve by the control of the control means;
The control means controls the valve opening timing of the intake valve and the exhaust valve to be longer in the low speed and high load operation region of the engine than in the high speed and high load operation region. Spark ignition gasoline engine. In the spark ignition gasoline engine according to any one of claims 1 to 7,
The exhaust pipe is
An independent exhaust pipe connected to the exhaust port for each cylinder;
A spark-ignited gasoline engine characterized by having a collecting portion that communicates downstream ends of independent exhaust pipes provided in cylinders whose combustion order is not adjacent to each other on the upstream side of the catalyst unit . The spark ignition gasoline engine according to claim 8 ,
The exhaust pipe has a plurality of collecting portions, and each collecting portion is gathered upstream of the catalyst unit .

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