JP2011153562A - Control device of spark ignition engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To enhance fuel consumption performance more effectively while avoiding misfire, etc. <P>SOLUTION: Within a specific load range of an engine, compression self-ignition combustion is performed in a low rotation range A of the engine, spark ignition combustion is performed with ignition of an air-fuel mixture by an ignition plug 16 in high rotation range B1 of the engine, in the low rotation range A in which the compression self-ignition combustion is performed, the more a load becomes small, the more the effective compression ratio ε' is raised, and in the high rotation range B1 in which the spark ignition combustion is performed, the more a load becomes small, the more the effective compression ratio ε' is lowered. <P>COPYRIGHT: (C)2011,JPO&INPIT

Description

  The present invention relates to a control device for a spark ignition engine having an ignition plug to which a fuel mainly composed of gasoline is supplied.

  Conventionally, in a spark-ignition gasoline engine having an ignition plug, depending on the operating region, compression self-ignition combustion in which the air-fuel mixture in the combustion chamber self-ignites and spark ignition combustion in which the air-fuel mixture is forcibly burned by spark ignition by the ignition plug And is known to do. 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 higher thermal efficiency is obtained compared to combustion by spark ignition. However, this compression self-ignition combustion has a problem that the combustion controllability deteriorates in a predetermined operating region (that is, premature ignition, knocking or misfire is likely to occur). Therefore, in such a region, it is necessary to perform forced combustion controlled by spark ignition, and compression self-ignition combustion and spark ignition combustion are selectively used depending on the operation region.

  For example, Patent Document 1 discloses an apparatus that performs compression self-ignition combustion in a low rotation / low load region and performs spark ignition combustion in other regions, specifically, a low rotation / high load region and a high rotation region. ing. With this system, compression self-ignition combustion is performed accurately in the low-rotation low-load region where combustion is easy to stabilize, and the thermal efficiency in the low-rotation low-load region is improved to improve fuel efficiency. I am trying. Further, in this device, the engine output is increased by setting the compression ratio to 13 or more in the high load region, while the pumping loss is suppressed by setting the compression ratio to a predetermined value of 13 or less in the low load region where compression self-ignition is performed. We are trying to improve performance.

JP 2007-292060 A

  In the device disclosed in Patent Document 1, in the low load region, compression self-ignition combustion is performed on the low rotation side and spark ignition combustion is performed on the high rotation side. In the region, the compression ratio is set to a predetermined value of 13 or less. Here, in the compression self-ignition combustion, there is a problem that combustion stability deteriorates when the compression ratio decreases. Therefore, if the compression ratio is simply kept small in the low load region as in this device, misfire or the like may occur during compression self-ignition combustion.

  This invention is made | formed in view of the above situations, and it aims at improving a fuel consumption performance more effectively, improving combustion stability.

  In order to solve the above-described problems, the present invention provides a control device for a spark ignition engine having an ignition plug capable of igniting an air-fuel mixture in a combustion chamber. Control means for controlling each part of the engine so that spark ignition combustion is performed by ignition of the air-fuel mixture by the spark plug in the high speed region of the engine while performing compression self-ignition combustion in the rotation region. Increases the effective compression ratio of the engine as the load decreases in the low rotation region where the compression self-ignition combustion is performed, while the effective compression ratio of the engine decreases as the load decreases in the high rotation region where the spark ignition combustion is performed. A control device for a spark ignition type engine is provided.

  According to this device, compression self-ignition combustion is performed in the low rotation region, while spark ignition combustion is performed in the high rotation region, and low rotation while reliably avoiding misfire due to compression self-ignition combustion in the high rotation region. Fuel efficiency can be improved by reliably realizing compression self-ignition combustion in the region. In addition, even in the same load region, the control of the effective compression ratio with respect to the load is varied according to the combustion mode, and the fuel efficiency can be improved in each combustion mode. In other words, the effective compression ratio is increased as the load decreases in the low rotation range where compression self-ignition combustion is performed, and the deterioration of the combustion stability accompanying the decrease in load is suppressed to achieve proper compression self-ignition combustion. The fuel consumption performance can be improved by the above, and the effective compression ratio is lowered along with the load reduction in the high rotation region where the spark ignition combustion is performed, and the fuel consumption performance can be improved by reducing the pumping loss.

  Here, in order to realize proper compression self-ignition combustion, relatively high values are required as the temperature and pressure during compression. Therefore, it is preferable that the geometric compression ratio of the engine body is set to 16 or more (claim 2).

  In the present invention, it is preferable that the control means sets the excess air ratio λ with respect to the stoichiometric air-fuel ratio in the combustion chamber to λ ≧ 2 over the entire operation region where the compression self-ignition combustion is performed. ).

  In this way, by performing the compression self-ignition combustion under a significantly lean air-fuel ratio of λ ≧ 2, the amount of NOx generated by the combustion can be effectively reduced, and the exhaust gas performance can be improved.

  Further, in the present invention, the control means sets the effective compression ratio of the engine in the low rotation region in which the compression self-ignition combustion is performed in at least a part of the specific load region to the spark ignition at the same load. In the transition from the high rotation region where the spark ignition combustion is performed to the low rotation region where the compression self-ignition combustion is performed, after the transition Preferably, the ignition of the air-fuel mixture by the spark plug is continued for a predetermined period of time.

  In this way, the effective compression ratio at the time of carrying out the compression self-ignition combustion can be set to a high value, and proper combustion can be realized more reliably, and effective from the spark-down combustion performed at a low effective compression ratio. When shifting to compression self-ignition combustion performed in a state where the compression ratio is high, it is possible to more reliably avoid occurrence of misfire or the like. That is, at the time of the transition, the effective compression ratio is not switched immediately and the effective compression ratio is not sufficiently increased, and there is a possibility that proper compression self-ignition combustion is difficult to achieve. Combustion is promoted by sustaining ignition by the spark plug and applying energy to the air-fuel mixture for a predetermined period later, and combustion stability can be improved.

  The predetermined period is preferably a period until the effective compression ratio of the engine reaches a value set for the low rotation region in which the compression self-ignition combustion is performed. That is, after the control means shifts from the high rotation region where the spark ignition combustion is performed to the low rotation region where the compression self ignition combustion is performed, the effective compression ratio of the engine is the low rotation region where the compression self ignition combustion is performed. It is preferable that the ignition of the air-fuel mixture by the spark plug is continued for a period until reaching a value set for (Claim 5).

  In this way, it is possible to more surely realize proper compression self-ignition combustion after the transition.

  As described above, according to the present invention, the fuel efficiency performance of the engine can be improved more effectively while avoiding misfire and the like by appropriately controlling the effective compression ratio of the engine according to the combustion mode.

It is a graph which shows the relationship between the excess air ratio (lambda) with respect to a theoretical air fuel ratio, and the production amount of NOx. It is a figure for demonstrating a mode that a fuel and oxygen react chemically and compression auto-ignition combustion occurs. It is a figure showing typically a general reciprocating gasoline engine. It is the graph which computed the conditions of the compression end temperature and the compression end pressure for making an air-fuel mixture self-ignite by MBT when an engine speed is 1000 rpm. It is a figure which shows the whole structure of the control apparatus of the spark ignition type engine concerning embodiment of this invention. It is a figure which shows an example of the lift characteristic of the intake valve set in order to control the effective compression ratio concerning embodiment of this invention. It is a figure which shows the control map concerning embodiment of this invention. It is a figure which shows the change of the effective compression ratio according to the engine load in a HCCI combustion area | region and a 2nd SI combustion area | region. It is a figure which shows the change of the effective compression ratio according to the engine load in a 1st SI combustion area | region and a 2nd SI combustion area | region. It is a figure which shows the change of the effective compression ratio at the time of combustion area transfer, and an ignition signal.

<Research up to the present invention>
(1-1) Consideration on air-fuel ratio In order to improve the thermal efficiency of the engine and improve the fuel efficiency, the air-fuel mixture may be burned (lean combustion) under an air-fuel ratio leaner than the stoichiometric air-fuel ratio. If the air-fuel ratio is made lean and the amount of air in the mixture is excessive with respect to the fuel (gasoline), the combustion temperature of the mixture will decrease, reducing engine exhaust loss and cooling loss, improving thermal efficiency To do.

  A gasoline engine (lean burn gasoline engine) that performs lean combustion as described above has already been developed. For example, a gasoline engine in which the air-fuel ratio of the air-fuel mixture is set to about 20 with respect to the theoretical air-fuel ratio, which is usually about 14.7, has been put into practical use. However, when the air-fuel ratio is about 20, there is a problem that not only a significant improvement in thermal efficiency can be expected, but also the emission performance deteriorates. In other words, the engine exhaust passage is generally provided with a three-way catalyst capable of simultaneously purifying HC, CO, and NOx contained in the exhaust gas. This three-way catalyst is used when the air-fuel ratio is the stoichiometric air-fuel ratio. The maximum performance is exhibited, and when the air-fuel ratio is leaned to about 20 as described above, the NOx purification performance is extremely lowered.

  Here, for example, by providing a NOx catalyst for storing and reducing NOx separately from the three-way catalyst, it is possible to purify NOx to some extent. However, even when the NOx catalyst is provided, when the air-fuel ratio is about 20, there is a possibility that the exhaust gas regulation that makes the exhaust amount of NOx stricter year by year cannot be sufficiently met.

  Therefore, the inventor of the present application thought to reduce the NOx amount (raw NOx amount) itself generated by combustion by making the air-fuel ratio much leaner than the stoichiometric air-fuel ratio. In other words, by making the air-fuel ratio significantly lean, the combustion temperature of the air-fuel mixture is made lower than the NOx generation temperature (the temperature at which NOx is actively generated), thereby reducing the amount of NOx generated did.

  FIG. 1 is a graph showing the relationship between the excess air ratio λ with respect to the stoichiometric air-fuel ratio and the amount of NOx produced (the amount of raw NOx produced by combustion). In this graph, the value Y on the vertical axis indicates the amount of NOx produced that sufficiently satisfies the predetermined exhaust gas regulation if the NOx catalyst is provided, and the value X indicates the above exhaust gas regulation without the provision of the NOx catalyst. The production amount of NOx that is sufficiently satisfactory is shown.

  As can be understood from FIG. 1, if the excess air ratio λ with respect to the stoichiometric air-fuel ratio is λ ≧ 2.4 (the air-fuel ratio is about 35 or more), the amount of raw NOx generated by combustion becomes the reference value X or less, Even without the NOx catalyst, exhaust gas regulations can be fully satisfied. Further, when the excess air ratio λ is in the range of 2 to 2.4, the NOx generation amount can be suppressed to be smaller than the reference value Y, and the regulation can be sufficiently satisfied by providing the NOx catalyst. .

  From the above, it can be seen that the excess air ratio λ with respect to the stoichiometric air-fuel ratio should be set to λ ≧ 2, more preferably λ ≧ 2.4 in order to clear the problem of NOx emission. If λ is set to 2.4 or more, the NOx catalyst can be omitted, which is advantageous in terms of cost as compared to when λ is 2 to 2.4.

  However, when the air-fuel mixture is burned under an ultra-lean air-fuel ratio of λ ≧ 2 (or 2.4), the flame propagation speed is greatly reduced as compared with the case where it is burned under the stoichiometric air-fuel ratio. Therefore, when combustion by spark ignition (hereinafter sometimes referred to as SI combustion) similar to that of a conventional gasoline engine is applied, there is a problem that the piston descends before the flame sufficiently propagates and misfires. On the other hand, when compression self-ignition combustion (hereinafter sometimes referred to as HCCI combustion) in which the air-fuel mixture self-ignites simultaneously and frequently is performed, λ ≧≧ regardless of the decrease in the flame propagation speed as described above. It was found that there is a possibility that proper combustion can be performed even under an air-fuel ratio of 2. Therefore, the inventor of the present application paid attention to this compression self-ignition combustion and further studied as follows.

(1-2) Consideration about HCCI combustion As shown in FIG. 2, HCCI combustion in a gasoline engine is a phenomenon in which fuel (gasoline) and oxygen voluntarily undergo a chemical reaction, generating thermal energy while generating water and carbon dioxide. Is generated. Whether or not the chemical reaction between the fuel and oxygen occurs is determined by the temperature, pressure, and high-temperature and high-pressure time of the air-fuel mixture (the time during which the air-fuel mixture is exposed to high temperature and high pressure). Specifically, the higher the temperature, the faster the speed of the molecules in the gas mixture, and the higher the pressure (that is, the higher the molecular density), the more frequent the collision of the molecules in the gas mixture. The higher the pressure, the easier the chemical reaction takes place. A chemical reaction, that is, combustion starts when a time during which the temperature and pressure are high (high temperature and high pressure time) continues to some extent, and thereafter, the combustion of the air-fuel mixture is completed by the chain reaction of the chemical reaction.

  As described above, in the HCCI combustion, the air-fuel mixture is self-ignited simultaneously and frequently, so that the combustion of the air-fuel mixture is completed in a shorter period of time than the combustion at the time of spark ignition, that is, SI combustion. However, as described above, the HCCI combustion starts when the high temperature and high pressure time continues to some extent. Therefore, in the operating region where the engine speed is high, this high temperature and high pressure time cannot be sufficiently secured, and misfire may occur due to the fact that HCCI combustion does not start. Therefore, in such an operation region where the engine speed is high, it is necessary to reliably avoid misfire by not performing HCCI combustion but performing SI combustion under an air excess ratio λ = 1, that is, an air-fuel ratio in the vicinity of the theoretical air-fuel ratio. There is. On the other hand, since a high temperature and high pressure time can be sufficiently secured in the low rotation region, HCCI combustion is performed under an air-fuel ratio with an excess air ratio λ ≧ 2, thereby improving fuel efficiency and reducing NOx.

  In performing the HCCI combustion, it is desirable that the air-fuel mixture is self-ignited at an optimum ignition timing (hereinafter referred to as MBT) at which the maximum torque is obtained. Therefore, the inventor of the present application considered the conditions of the in-cylinder temperature T and the in-cylinder pressure P for self-igniting the air-fuel mixture with the MBT in the gasoline engine as shown in FIG. FIG. 3 schematically shows a general reciprocating gasoline engine composed of pistons, cylinders and the like. The MBT is generally between the timing near the compression top dead center (at the time of high load) and ATDC (after compression top dead center) 3 ° CA (at the time of low load), although it varies depending on the engine load.

  FIG. 4 shows the compression end temperature (in-cylinder temperature immediately before MBT) Tx and the compression end pressure at which the air-fuel mixture self-ignites with MBT when the engine excess speed is 1000 rpm at the excess air ratio λ = 2.4 with respect to the stoichiometric air-fuel ratio. (Cylinder pressure immediately before MBT) Px is a graph calculated based on elementary reaction calculation and state equation. In order to auto-ignite the air-fuel mixture under a lean air-fuel ratio of λ = 2.4, it is necessary to increase the temperature and pressure of the air-fuel mixture at a higher compression ratio than a general reciprocating gasoline engine. In terms of thermal efficiency, a higher compression ratio is advantageous. Therefore, in calculating this graph, the displacement per engine cylinder was 500 cc, and the geometric compression ratio was 18.

  In the graph of FIG. 4, the line L1 is a line connecting the values of the compression end temperature Tx and the compression end pressure Px necessary for the air-fuel mixture to self-ignite with MBT, and each value Tx, Px is on the line L1. If it is, the self-ignition timing is MBT. In addition, the compression end temperature Tx and the compression end pressure Px in this figure refer to the in-cylinder temperature and pressure at the compression top dead center.

  In the graph of FIG. 4, lines M1, M2, M3 and M4 indicate the relationship between temperature and pressure immediately before the MBT (compression end) when the amount of fresh air introduced into the cylinder, that is, constant load, The load decreases in order from M1. That is, since the stroke volume immediately before the MBT (compression end) is always constant when the geometric compression ratio is constant 18, the fresh air amount (load) is proportional to the compression end pressure Px according to the state equation. It is inversely proportional to Tx. For this reason, a plurality of lines M1 to M4 having different inclinations can be defined for each load. In this figure, the load of each line M1 to M4 is represented by the indicated mean effective pressure (IMEP), M1 is full load IMEP = 1300 kPa, M2 is IMEP = 900 kPa, M3 is 1/3 load IMEP = 500 kPa, M4 is unloaded IMEP = 200 kPa.

  As shown in the graph of FIG. 4, the condition (line L1) in which the air-fuel mixture self-ignites with MBT is inclined downward to the right. The lower the compression end pressure Px, the higher the compression end temperature Tx needs to be. That is, in order to cause the above-described chemical reaction, the compression end temperature Tx must be increased to increase the molecular velocity by the amount that the compression end pressure Px has decreased and the collision frequency of molecules has decreased.

  Note that when the compression end temperature Tx and the compression end pressure Px are deviated to a higher temperature / high pressure side (upper right side of the graph) than the line L1, the self-ignition timing is earlier than the MBT, whereas the above condition is the line L1. If the temperature falls outside the lower temperature / low pressure side (lower left side of the graph), the self-ignition timing becomes later than MBT. If the line L1 deviates greatly to the high temperature / high pressure side or the low temperature / low pressure side, pre-ignition or knocking occurs on the high temperature / high pressure side, and misfire occurs on the low temperature / low pressure side.

  The intersection of the line L1 showing the conditions for self-ignition with MBT and the lines M1 to M4 showing the relationship between pressure and temperature at each load is the temperature-pressure condition for realizing self-ignition with MBT at each load. is there. As shown in the graph of FIG. 4, it is necessary to lower the compression end pressure Px and increase the compression end temperature Tx in order to cause the MBT to self-ignite as the load decreases.

  As a specific method for increasing the compression end temperature Tx in response to a decrease in load, for example, there is a method of heating the temperature of fresh air. However, it is not realistic to heat or cool fresh air over a wide temperature range as shown in FIG. 4 due to problems such as cost and controllability.

(1-3) Solution Therefore, the inventors of the present application focused on the point that the compression end temperature Tx can be changed by changing the effective compression ratio of the engine. That is, it was considered that the effective compression ratio of the engine is increased as the load is reduced, thereby increasing the compression end temperature Tx and realizing HCCI combustion in MBT.

  In order to change the effective compression ratio of the engine, it is only necessary to change the timing when the compression starts substantially by changing the closing timing of the intake valve. The compression end temperature Tx can be changed without significantly changing from the above.

<Embodiment of the present invention>
(2-1)
Next, a spark ignition engine control apparatus according to an embodiment of the present invention devised based on the basic theory as described above will be described with reference to the drawings.

  FIG. 5 is a diagram showing the overall configuration of the control device for the spark ignition engine according to the embodiment of the present invention. The engine shown in the figure includes a cylinder block 3 having a plurality of cylinders 2 (only one of which is shown in the figure) arranged in a direction perpendicular to the paper surface, and a cylinder arranged on the cylinder block 3. It has an engine body 1 composed of a multi-cylinder gasoline engine equipped with a head 4. 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. It may be.

  A piston 5 is inserted into each cylinder 2 of the engine body 1 so as to be able to reciprocate. The piston 5 is connected to the crankshaft 7 via a connecting rod 8, and the crankshaft 7 rotates around its central axis in accordance with the reciprocating motion of the piston 5.

  A combustion chamber 6 is formed above the piston 3. 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 intake valve 11 and the exhaust valve 12 are driven to open and close in conjunction with the rotation of the crankshaft 7 by a valve mechanism 13 including a pair of camshafts (not shown) disposed in the cylinder head 4.

  Each valve operating mechanism 13 for the intake valve 11 and the exhaust valve 12 includes a VVL 14 and a VVT 15, respectively. VVL14 is an abbreviation for variable valve lift mechanism, and variably sets the lift amount (valve opening amount) of the intake and exhaust valves 11 and 12. The VVT 15 is an abbreviation for a variable valve timing mechanism, and variably sets the opening / closing timing (phase angle) of the intake and exhaust valves 11 and 12. The VVL 14 and VVT 15 are known in various forms already in practical use, and detailed description thereof is omitted here, but those disclosed in, for example, Japanese Patent Application Laid-Open No. 2007-85241 are employed. It is possible.

  The cylinder head 4 of the engine body 1 is provided with a spark plug 16 so as to face the combustion chamber 6 of each cylinder 2 from above. The spark plug 16 is electrically connected to an ignition circuit 17 provided in the upper part of the cylinder head 4, and an ignition spark is discharged from the spark plug 16 at a predetermined timing in response to power supply by the ignition circuit 17. Is done.

  The cylinder head 4 is provided with an injector 18 so that the combustion chamber 6 faces the intake side. In the intake stroke of the engine or the like, fuel (fuel mainly composed of gasoline) is directly injected from the injector 18 into the combustion chamber 6 and the injected fuel is mixed with air. Is produced.

  In the engine main body 1 configured as described above, the geometric compression ratio determined based on the volume of the combustion chamber 6 when the piston 5 is at the top dead center and the stroke volume of the piston 5 is set to 18. .

  An intake passage 20 and an exhaust passage 21 are connected to the intake port 9 and the exhaust port 10 of the engine body 1, respectively. That is, combustion air (fresh air) is supplied to the combustion chamber 6 through the intake passage 20, and burned gas (exhaust gas) generated in the combustion chamber 6 is discharged to the outside through the exhaust passage 21. It is like that.

  The supercharger 25 includes a compressor 26 provided in the intake passage 20, a turbine 27 provided in the exhaust passage 21, and a connecting shaft 28 that connects the compressor 26 and the turbine 27. When the turbine 27 rotates in response to the energy of the exhaust gas, the compressor 26 rotates at a high speed in conjunction with this, so that fresh air passing through the intake passage 20 is pressurized and pumped to the combustion chamber 6. Is done.

  The exhaust passage 21 is provided with a bypass passage 33 for bypassing the turbine 27 and an electric waste gate valve 34 for opening and closing the bypass passage 33. Then, the waste gate valve 34 is operated to open and close the bypass passage 33, so that the exhaust gas flows into the turbine 27 and the turbine 27 is driven to rotate, and the exhaust gas bypasses the turbine 27 and bypasses the turbine 27. It can be switched to a state where the rotation of the motor is stopped.

  The exhaust passage 21 is provided with a catalytic converter 32 for purifying exhaust gas. A three-way catalyst is built in the catalytic converter 32, and harmful components in the exhaust gas passing through the exhaust passage 21 are purified by the action of the three-way catalyst.

(2-2) Control system The engine configured as described above is provided with an ECU (Engine Control Unit) 40 including a conventionally known CPU, memory, and the like as a control means for comprehensively controlling its operation. Yes.

  The ECU 40 is electrically connected to sensors provided in each part of the engine. Specifically, the ECU 40 includes an engine speed sensor 51 that detects the rotational speed of the crankshaft 7, that is, the engine speed, an airflow sensor 52 that detects the amount of fresh air passing through the intake passage 20, and a driver. And an accelerator opening sensor 53 that detects an operation amount (accelerator opening) of an accelerator pedal (not shown) that is depressed by the operation of the accelerator pedal is electrically connected. Each is input to the ECU 40.

  The ECU 40 is also electrically connected to the VVL 14, VVT 15, ignition circuit 17, injector 18, throttle valve 22, and waste gate valve 34, and outputs control signals for driving to these devices. It is configured.

  The more specific functions of the ECU 40 will be described. The ECU 40 includes valve control means 41, supercharging control means 42, injector control means 43, and ignition control means 44 as main functional elements. Yes.

  The valve control means 41 drives the VVL 14 and VVT 15 to variably set the lift characteristics (open / close timing and lift amount) of the intake and exhaust valves 11 and 12. More specifically, the valve control means 41 has a function of controlling the effective compression ratio of the engine by changing the lift characteristics of the intake and exhaust valves 11 and 12 as described above.

  When the valve control means 41 controls the effective compression ratio of the engine, for example, the lift characteristic of the intake valve 11 is changed in the manner shown in FIG. That is, as shown by the solid line waveform in FIG. 6, when the intake valve 11 is closed near the retarded side of the intake bottom dead center (timing slightly past the intake bottom dead center), the effective compression ratio is geometric. This is consistent with the dynamic compression ratio (18 in this embodiment). On the other hand, as shown by the broken line waveform in FIG. 6, in the state where the intake valve 11 is set later than the intake bottom dead center is further closed on the retard side, the substantial start timing of the compression stroke is delayed. As a result, the substantial compression ratio (effective compression ratio) of the engine decreases. The valve control means 41 variably sets the effective compression ratio of the engine by increasing or decreasing an amount (retard amount) that delays the closing timing of the intake valve 11.

  FIG. 6 illustrates a case where the VVT 15 for the intake valve 11 is operated and the operation timing (opening and closing timing) of the intake valve 11 is shifted to the retard side as shown by the broken line in FIG. Yes. As described above, when only the VVT 15 is operated, not only the closing timing of the intake valve 11 but also the opening timing of the intake valve 11 is changed. Therefore, in the present embodiment, the VVL 14 is operated together with the VVT 15 so that the opening timing of the intake valve 11 is not changed.

  The injector control means 43 controls the injection timing and injection amount (injection time) of fuel injected from the injector 18 into the combustion chamber 6. More specifically, the injector control means 43 calculates a target injection amount of fuel for obtaining a predetermined air-fuel ratio based on information such as the intake air amount (fresh air amount) input from the air flow sensor 52. To do. The in-cylinder air-fuel ratio is controlled by opening the injector 18 for a time corresponding to the target injection amount. In this embodiment, the injector control means 43 maintains the excess air ratio λ with respect to the stoichiometric air-fuel ratio at λ = 2.4 in the HCCI combustion region described later, while λ = 1 in the SI combustion region described later. The fuel injection amount from the injector 18 is controlled so as to be maintained at 0.0. Here, in this embodiment, in order to ensure a sufficient mixing time of the fuel and air, the injector control means 43 controls the injection timing of the injector 18 so as to inject fuel during the intake stroke.

  The supercharging control means 42 controls the supercharger 25 so as to obtain an appropriate supercharging pressure, that is, a fresh air amount according to the operating state, by opening and closing the waste gate valve 34.

  The ignition control means 44 controls the timing of spark discharge to the air-fuel mixture by the spark plug 16 by controlling the power supply from the ignition circuit 17 to the spark plug 16.

(2-3) Specific Example of Control Next, how the engine is controlled according to the load and the rotational speed by the ECU 40 configured as described above will be specifically described.

  FIG. 7 is a diagram showing a control map referred to when the ECU 40 controls the engine. In FIG. 7, the HCCI region A set in the low load / low rotation region where the engine load is low and the engine speed is low is an operation region in which HCCI combustion is performed, and the first set in the low load / high rotation region. The first SI region B1 and the second SI region B2 set in a high load region with a high engine load are operating regions in which SI combustion is performed. More specifically, a region where the engine speed is low in the region where the engine load is equal to or less than the predetermined value Q1 is set as the HCCI region A, and is the same load region as the HCCI region A and the engine speed is higher than the HCCI region A. The high area is set as the first SI area B1. And the area | region where engine load is higher than these HCCI area | region A and 1st SI area | region B1 is set to 2nd SI area | region B2.

  As described above, a high temperature and high pressure time cannot be sufficiently secured in a high rotation region where the engine speed is high, and misfire occurs in HCCI combustion. Therefore, in this engine, even in the same load region, HCCI combustion is performed only in the low rotation region A and SI combustion is performed in the high rotation region B1, and misfiring is performed while improving fuel efficiency performance by HCCI combustion. Avoid more reliably.

  As described above, in order to perform HCCI combustion with MBT, it is necessary to decrease the compression end pressure Px and increase the compression end temperature Tx as the load decreases. In other words, it is necessary to increase the compression end pressure Px while lowering the compression end temperature Tx as the load increases. Here, as the load increases, the compression end pressure Px increases to some extent as the amount of fresh air increases. However, as shown in the graph of FIG. 5, in the high load region of the predetermined value Q1 or more (in the graph of FIG. 5, the load indicated by the line M2 is a high load region of about IMEP 900 or more) The value of the required compression end pressure Px is much higher than the pressure obtained by the supercharging performance of the conventional supercharger. For this reason, it is necessary to change the structure of the supercharger in order to cause the MBT to self-ignite in such a high load region. Therefore, in this engine, in the high load region, high engine output is ensured without changing the conventional system configuration by stopping HCCI combustion and performing SI combustion.

  The injector control means 43, the ignition control means 44, and the valve control means 41 are combusted so that HCCI combustion and SI combustion are properly performed in accordance with the engine operating region, that is, the engine speed and load. The excess air ratio in the room, the ignition timing, etc. and the effective compression ratio ε ′ of the engine are controlled. Further, the supercharging control means 43 opens and closes the waste gate valve 34 to appropriately pressurize fresh air, and introduce a necessary amount of fresh air into each combustion chamber at each load.

  Specifically, in the HCCI combustion region A set in the low load and low rotation region, the injector control means 43 is an amount that provides an excess air ratio λ = 2.4 with respect to the amount of fresh air introduced into the combustion chamber. The injector 18 is driven so as to inject this fuel. On the other hand, in the first SI combustion region B1 and the second combustion region B2, the injector 18 is set so that an amount of fuel with an excess air ratio λ = 1.0 is injected with respect to the amount of fresh air introduced into the combustion chamber. To drive.

  The ignition control means 44 stops the ignition of the air-fuel mixture by the spark plug 16 in the HCCI combustion region A. On the other hand, the air-fuel mixture is ignited by the spark plug 16 in the first SI combustion region B1 and the second combustion region B2.

  In the HCCI combustion region A, the valve control means 41 controls the VVT 15 and VVL 14 so that the effective compression ratio ε ′ increases as the load decreases. On the other hand, in the SI combustion region B1 set in the low load / rotation region, the VVT 15 and the VVL 14 are controlled so that the effective compression ratio ε ′ decreases as the load decreases.

  As described above, in order to realize HCCI combustion in MBT, it is necessary to increase the effective compression ratio ε ′ as the load decreases. Therefore, in the HCCI combustion region A, the valve control means 41 controls the VVT 15 and VVL 14 so that the effective compression ratio ε ′ increases as the load decreases. Specifically, from the timing when the closing timing of the intake valve 11 is significantly retarded from the intake bottom dead center, the closing timing of the intake valve 11 is advanced as the load decreases so as to approach the intake bottom dead center. Controls VVT 15 and the like. FIG. 8 shows a change in the effective compression ratio ε ′ along the line indicated by U1 in FIG. In the present embodiment, as shown in FIG. 8, the effective compression ratio ε ′ is controlled to about 14 with the closing timing of the intake valve 11 being 65 ° C. when the engine load is in the vicinity of IMEP900. Then, as the load decreases, the closing timing of the intake valve 11 is advanced, and in the vicinity of no load, the closing timing is set to 0 ° C., and the effective compression ratio ε ′ is controlled to about 18 which is the same as the geometric compression ratio. To do.

  As described above, in this engine, the effective compression ratio ε ′ is increased as the load is reduced, so that the compression end temperature Tx of the engine is increased and HCCI combustion in MBT is realized. In particular, the misfire caused by the low compression end temperature Tx of the engine is surely avoided, and the improvement of the fuel consumption performance accompanying the realization of HCCI combustion is realized.

  Here, when the effective compression ratio ε ′ is increased, the amount of work required for intake and compression increases and so-called pumping loss increases. Therefore, in the first SI combustion region B1, as in the HCCI combustion region A, if the effective compression ratio ε ′ is increased as the load decreases, the fuel efficiency deteriorates due to an increase in pumping loss. Of course, even in the HCCI combustion, the fuel efficiency deteriorates with the increase in the effective compression ratio ε ', but the effect of improving the fuel efficiency by realizing the HCCI combustion in the MBT is high, and the fuel efficiency can be improved as a whole. However, in SI combustion, the effect of the pumping loss on the fuel consumption performance is large, and the fuel consumption performance deteriorates due to an increase in the effective compression ratio ε '. In particular, in the low load region where the engine output is small, the influence of the pumping loss is large.

  Therefore, in this engine, in the first SI combustion region B1, the overall effective compression ratio ε ′ is kept lower than that in the HCCI combustion region A, and the effective compression ratio ε ′ is lowered as the load decreases to reduce the pumping loss. Improves fuel efficiency by suppressing it. FIG. 9 shows a change in the effective compression ratio ε ′ along the line indicated by U2 in FIG. In this embodiment, as shown in FIG. 9, when the engine load is around IMEP 900 kPa, the effective compression ratio ε ′ is controlled to about 14 while the closing timing of the intake valve 11 is set to 65 ° C. as in the HCCI combustion. As the load decreases, the closing timing is retarded, and in the vicinity of no load, the closing timing is set to 90 ° C. and the effective compression ratio ε ′ is reduced to about 10.5.

  Thus, in this engine, even in the same low-load region, in the low-speed region A where HCCI combustion is performed, the effective compression ratio ε ′ is increased as the load is reduced, and HCCI combustion by MBT is realized to improve fuel efficiency. In the high rotation range B1 where SI combustion is performed, the effective compression ratio ε ′ is reduced as the load decreases to reduce the pumping loss, thereby improving the fuel consumption performance, thereby improving the fuel consumption performance in the entire low load range. Yes.

  In the second SI combustion region B2 in the high load region, control is performed so that the effective compression ratio ε ′ = 14 is constant. Here, if the effective compression ratio ε ′ is increased, the engine output can be increased. However, if ε ′ is set to a high value of 14 or more, knocking may occur in SI combustion. Therefore, in this embodiment, the effective compression ratio ε ′ = 14 is constant in order to ensure engine output while avoiding knocking in the high load region B2.

  Next, the control performed at the time of transfer from the HCCI combustion region A to the first SI combustion region B1 will be described.

  As described above, the effective compression ratio ε ′ set in the HCCI combustion region A on the low rotation side is set to a higher value than the effective compression ratio ε ′ set in the first SI combustion region B1 on the high rotation side. ing. Therefore, at the time of transition from the first SI combustion region B1 to the HCCI combustion region A, it is necessary to immediately advance the closing timing of the intake valve 11 by the valve control means 41 in order to change the effective compression ratio ε ′ to the higher side. is there.

  However, there is a limit to the driving speed of the VVT 15 that is driven to change the closing timing of the intake valve 11, and, as shown in FIG. The value shifts from the value on the 1SI combustion region B1 side to the high effective compression ratio ε ′ on the HCCI combustion region A side. That is, immediately after the transition, the effective compression ratio ε ′ is not sufficiently increased, and a compression end pressure Px sufficient to realize HCCI combustion is not ensured. Therefore, in this state, there is a risk of misfire if the ignition of the air-fuel mixture by the spark plug 16 stops as indicated by the broken line in FIG.

  Therefore, in this engine, at the time of transition from the first SI combustion region B1 to the HCCI combustion region A, misfire is avoided by igniting the air-fuel mixture by the spark plug 16 and promoting combustion for a predetermined period.

  In the present embodiment, as shown by the solid line in FIG. 10, until the effective compression ratio ε ′ reaches the effective compression ratio ε ′ set for the HCCI combustion, that is, the closing timing of the intake valve 11 is for HCCI combustion. Until the set timing is reached, the ignition signal is output and the ignition of the air-fuel mixture by the spark plug 16 is continued. Thereby, the effective compression ratio ε 'is set to a different value between the HCCI combustion region A and the first SI combustion region B1, and the misfire at the time of transition can be avoided while improving the fuel efficiency performance in these regions as a whole. Here, the value of the effective compression ratio ε ′, that is, the actual closing timing of the intake valve 11 can be calculated from the value of the cam angle sensor provided for detecting the phase of the camshaft that drives the intake valve 11. Good.

  As described above, in the engine of the present embodiment, HCCI combustion is performed in the low rotation region in the low load region, while SI combustion is performed in the high rotation region, and HCCI combustion can be reliably performed, While the effective compression ratio ε ′ is increased as the load decreases in the low rotation region A where HCCI combustion is performed, the effective compression ratio ε ′ is decreased as the load decreases in the high rotation region B1 where SI combustion is performed, thereby avoiding misfires and the like. Thus, it is possible to achieve proper HCCI combustion in the low rotation region A and reduce the pumping loss in the high rotation region B1, thereby improving the overall fuel efficiency.

  In addition, the effective compression ratio ε ′ of the HCCI combustion region A is made higher than that of the first SI combustion region B1, and misfire of the HCCI combustion is more reliably avoided, and the pumping loss in the first SI combustion region B1 is kept small. At the time of transition from the SI combustion region having a different effective compression ratio ε ′ to the HCCI combustion region, the ignition of the air-fuel mixture by the spark plug 16 is continued even after the transition to assist combustion by spark ignition. Misfires associated with transition can be reliably avoided.

  In the above-described embodiment, the load region in which the HCCI combustion and the SI combustion are made different depending on the rotation speed, that is, the HCCI combustion region A and the first SI combustion region B1, is set between the no-load region and the predetermined load region (IMEP 900 kPa). Although it was set, this is not restrictive. For example, HCCI combustion is performed up to the high load region by using a high supercharging device, etc., HCCI combustion is performed on the low rotation side over all load regions, and SI combustion is performed on the high rotation side You may make it do.

  In the above embodiment, by setting the excess air ratio λ with respect to the stoichiometric air-fuel ratio in the HCCI combustion region A to λ = 2.4, the NOx amount (raw NOx amount) generated by combustion itself is sufficiently reduced, Even if the NOx catalyst is omitted, exhaust gas regulations can be sufficiently satisfied. However, if it is possible to provide a NOx catalyst, the excess air ratio λ may be made smaller than 2.4. However, even if a NOx catalyst is provided, the excess air ratio λ should be set to at least λ ≧ 2 in order to sufficiently satisfy strict exhaust gas regulations expected in the future. As described with reference to FIG. 1, if the excess air ratio λ is 2 or more, it is possible to suppress the emission amount to a sufficient level by purifying the generated NOx with a NOx catalyst.

  In the above embodiment, the case where the excess air ratio λ = 1.0 with respect to the stoichiometric air-fuel ratio in the SI combustion regions B1 and B2 has been described. For example, the energy of the flame kernel is increased by performing multiple ignition or the like. If the misfire can be avoided by doing so, the excess air ratio λ may be 2 or more.

  Moreover, in the said embodiment, although the geometric compression ratio of the engine was set to 18, you may set to compression ratios other than this. However, in the graph of FIG. 4, a region W indicating a temperature / pressure range when fresh air at atmospheric pressure is compressed at a compression ratio of 16 to 18 is used to realize compression self-ignition combustion in the MBT. Since it is located on the line L1 indicating the condition, if the geometric compression ratio is set to at least 16 or more, the effective compression ratio ε ′ is appropriately changed with the compression ratio 16 being the maximum value, thereby implementing the above-described implementation. Similar to the embodiment, it is considered possible to perform appropriate compression self-ignition combustion at each load. For this reason, the geometric compression ratio is preferably set to at least 16 or more.

  Further, in the control for maintaining the spark ignition performed at the time of transition from the SI combustion region to the HCCI combustion region for a predetermined period, the predetermined period for maintaining the spark ignition is not limited to the above. That is, the predetermined period may be set to a predetermined time period instead of the period until the effective compression ratio ε ′ becomes the effective compression ratio ε ′ for HCCI combustion as described above.

6 Combustion chamber 11 Intake valve 16 Spark plug 18 Injector 40 ECU (control means)
ε 'effective compression ratio

Claims (5)

A control device for a spark ignition engine having an ignition plug capable of igniting an air-fuel mixture in a combustion chamber,
In a specific load region of the engine, compression auto-ignition combustion is performed in a low rotation region of the engine, while spark ignition combustion is performed in the high rotation region of the engine by ignition of an air-fuel mixture by the spark plug. Comprising control means for controlling each part;
The control means increases the effective compression ratio of the engine as the load decreases in the low rotation region where the compression self-ignition combustion is performed, while the effective compression of the engine decreases as the load decreases in the high rotation region where the spark ignition combustion is performed. A control device for a spark ignition engine characterized by lowering the ratio. A control device for a spark ignition engine according to claim 1,
A control device for a spark ignition engine, wherein the geometric compression ratio of the engine body is set to 16 or more. A control device for a spark ignition engine according to claim 1 or 2,
The control device for a spark ignition engine, wherein the control means sets an excess air ratio λ with respect to the stoichiometric air-fuel ratio in the combustion chamber to λ ≧ 2 in the entire operation region where the compression self-ignition combustion is performed. A control device for a spark ignition engine according to any one of claims 1 to 3,
The control means is configured to increase the effective compression ratio of the engine in the low-rotation region in which the compression self-ignition combustion is performed in at least a part of the specific load region, and to perform the spark ignition combustion in the same load. The ignition ratio is set higher than the effective compression ratio of the engine in the region, and the ignition is performed for a predetermined period after the transition from the high rotation region where the spark ignition combustion is performed to the low rotation region where the compression self-ignition combustion is performed. A control device for a spark ignition engine characterized in that ignition of an air-fuel mixture by a plug is sustained. A control device for a spark ignition engine according to claim 4,
After the transition from the high rotation region where the spark ignition combustion is performed to the low rotation region where the compression self-ignition combustion is performed, the control means changes the effective compression ratio of the engine to the low rotation region where the compression self-ignition combustion is performed. A spark ignition type engine control device, characterized in that ignition of the air-fuel mixture by the spark plug is continued for a period until a value set for the spark plug is reached.

Images