JP2011021553A - Method and device for controlling engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To perform compression self-ignition combustion under a lean air-fuel ratio in a wider operational area while preventing an excessive increase in cylinder internal pressure. <P>SOLUTION: In this method for controlling an engine, the air-fuel ratio of an air-fuel mixture is made lean in comparison with a theoretical air-fuel ratio in the whole area in the load direction of an HCCI area A where the compression self-ignition combustion is performed, and also when an engine load is increased to a predetermined load X or more in the HCCI area A, timing θig of self-ignition of the air-fuel mixture is retarded with respect to MBT as ignition timing with torque maximized. <P>COPYRIGHT: (C)2011,JPO&INPIT

Description

  The present invention relates to a method for controlling an engine in which compression self-ignition combustion is performed in which an air-fuel mixture is self-ignited by fuel injection from an injector and compression action of a piston in an HCCI region set in at least a part of the engine operation region. And device.

  Conventionally, as shown in Patent Document 1 below, in a spark ignition type gasoline engine having a plurality of cylinders, compression self-ignition combustion for self-igniting an air-fuel mixture is performed in a partial load operation region where the engine speed is equal to or lower than a predetermined speed. On the other hand, in other operation regions, the operation is performed by forced combustion by spark ignition.

  As described above, many proposals have been made to properly use compression self-ignition combustion and forced combustion by spark ignition according to the operating state of the engine. In other words, compression self-ignition combustion is combustion that is self-ignited simultaneously and frequently at various locations in the combustion chamber, and it is said that higher thermal efficiency can be obtained compared to combustion by spark ignition that has been conventionally performed, Since there is a problem in the combustion controllability in the high rotation / high load range, it is necessary to perform forced combustion controlled by spark ignition in such an operation range. Therefore, two types of combustion, combustion by compression self-ignition and combustion by spark ignition, are used properly according to the operating state.

JP 2007-154859 A

  By the way, in the field of engines, it is required to develop an engine with higher thermal efficiency and lower fuel consumption against the background of severe energy conditions in recent years. As measures to further increase the thermal efficiency, it is conceivable to perform compression auto-ignition combustion in a wider operating range and to make the air-fuel ratio of the air-fuel mixture lean, but trying to maintain a lean air-fuel ratio even in a high load range Then, it is necessary to send a large amount of air into the cylinder by supercharging or the like. However, when a large amount of air is fed into the cylinder, the in-cylinder pressure rises excessively, and there is a problem that the durability and reliability of the engine are impaired.

  The present invention has been made in view of the circumstances as described above, and can perform compression self-ignition combustion under a lean air-fuel ratio in a wider operating region while preventing an excessive increase in in-cylinder pressure. An object of the present invention is to provide a control method and a control device for an engine.

  In order to solve the above-mentioned problems, the present invention provides a compression self-ignition that self-ignites an air-fuel mixture by fuel injection from an injector and compression action of a piston in an HCCI region set in at least a part of an engine operation region. A method of controlling an engine in which ignition combustion is performed, wherein the air-fuel ratio of the air-fuel mixture is set to be leaner than the stoichiometric air-fuel ratio in the entire load direction of the HCCI region, and the engine load is set within the HCCI region. When the load is increased beyond the load, the self-ignition timing of the air-fuel mixture is retarded with respect to MBT, which is the ignition timing at which the torque becomes maximum (claim 1).

  According to the present invention, by making the air-fuel ratio lean throughout the load direction of the HCCI region, it is possible to effectively improve thermal efficiency and improve fuel efficiency. However, if the lean air-fuel ratio is continued up to the high load side, it is necessary to significantly increase the intake air amount as the load increases, resulting in an excessive increase in in-cylinder pressure that impairs engine durability and reliability. There is a risk of being. In order to deal with such a problem, in the present invention, since the ignition timing of the air-fuel mixture is retarded with respect to the MBT at a predetermined load or more, the situation where the in-cylinder pressure excessively rises in the high load region as described above is effectively prevented. As a result, there is an advantage that the compression self-ignition combustion under the lean air-fuel ratio can be continued to the higher load side and the thermal efficiency can be greatly improved.

  In the present invention, preferably, when the engine load increases to the predetermined load or more in the HCCI region, the fuel injection timing from the injector is retarded from the injection timing corresponding to the MBT (Claim 2).

  In this way, it is possible to effectively prevent an excessive increase in the in-cylinder pressure by appropriately retarding the ignition timing in the high load region with a relatively simple configuration that only changes the fuel injection timing from the injector. There are advantages.

  In the present invention, when the ignition assist for discharging sparks from the spark plug to assist the self-ignition of the air-fuel mixture is performed in the HCCI region, when the engine load exceeds the predetermined load in the region, the spark plug It is preferable to retard the ignition assist timing by the ignition assist timing corresponding to the MBT (claim 3).

  In this way, there is an advantage that the control for retarding the ignition timing at a predetermined load or more can be appropriately performed by changing the ignition assist timing by the spark plug.

  In the present invention, preferably, the combustion period of the compression self-ignition combustion in the HCCI region is made substantially constant regardless of whether or not the predetermined load or more.

  In this way, by maintaining parameters other than the ignition timing at appropriate values from the viewpoint of thermal efficiency, there is an advantage that an excessive increase in in-cylinder pressure can be effectively prevented while minimizing a decrease in thermal efficiency.

  In the present invention, preferably, the excess air ratio λ with respect to the stoichiometric air-fuel ratio is set to λ = 2 to 3 in the entire load direction of the HCCI region.

  In this way, both exhaust loss and cooling loss of the engine can be reduced in a balanced manner, and the thermal efficiency can be increased to a sufficient level within a feasible range.

  In the present invention, it is preferable to perform supercharging using a supercharger at least in a load range equal to or greater than the predetermined load (claim 6).

  In this way, it is possible to continue the lean air-fuel ratio to the high load region and further improve the thermal efficiency while sufficiently securing the engine output in the high load region by supercharging.

  Further, the present invention controls an engine in which compression self-ignition combustion is performed in which an air-fuel mixture is self-ignited by fuel injection from an injector and compression action of a piston in an HCCI region set in at least a part of the engine operation region. An injector control means for controlling the fuel injection operation from the injector, wherein the air-fuel ratio of the air-fuel mixture is set to be leaner than the stoichiometric air-fuel ratio in the entire load direction of the HCCI region. When the engine load increases above a predetermined load within the HCCI region, the injector control means executes a control for retarding the self-ignition timing of the air-fuel mixture with respect to MBT, which is the ignition timing at which the torque becomes maximum. (Claim 7).

  Even in the case of the present invention, it is possible to obtain the same operational effects as in the case of the control method described above.

  As described above, according to the engine control method and control apparatus of the present invention, it is possible to carry out compression self-ignition combustion under a lean air-fuel ratio in a wider operating region while preventing an excessive increase in in-cylinder pressure. There are advantages.

1 is a diagram illustrating an overall configuration of a spark ignition engine according to an embodiment of the present invention. It is sectional drawing of an engine main body. It is a block diagram which shows the control system of an engine. It is a figure which shows an example of the control map referred in order to control combustion of an engine. It is a figure which shows the example of control of an engine compression ratio, an air excess rate, an ignition timing, a combustion period, and a supercharging amount. It is a figure which shows the various loss factors which influence the thermal efficiency of an engine, and the various control parameters relevant to the loss factor. It is a figure which shows the driving | operation area | region of an engine, and its representative point. It is a figure which shows the calculation result of the illustration thermal efficiency in a representative point in relation to various parameters, such as a compression ratio and an excess air ratio. It is a figure which shows the calculation result of the exhaust loss in a representative point in relation to various parameters, such as a compression ratio and an excess air ratio. It is a figure which shows the calculation result of the cooling loss in a representative point in relation to various parameters, such as a compression ratio and an excess air ratio. It is a figure which shows the characteristic of the heat transfer coefficient at the time of setting an excess air ratio to several different value. It is a figure which shows the characteristic of the cooling loss integrated value at the time of setting an excess air ratio to several different value. It is a figure which shows each value of the in-cylinder pressure in a representative point, a pressure increase rate, and exhaust temperature with respect to a crank angle. It is a figure which shows the change of the largest in-cylinder pressure at the time of driving | running an engine to the maximum load, the largest pressure increase rate, and exhaust temperature. It is a figure which shows how the maximum in-cylinder pressure, the maximum pressure increase rate, and exhaust temperature change when various parameters, such as a compression ratio and an excess air ratio, are changed in the high load area of an engine. It is a figure which shows the change of the illustration thermal efficiency at the time of changing various parameters on the same conditions as FIG. It is the table | surface which put together the result of FIG. 15 and FIG. It is a figure which shows the example of desirable control obtained from the result of FIGS. 15-17, and the change of the illustration thermal efficiency based on it.

A. Embodiment (1) Basic Configuration of Engine FIG. 1 is a diagram showing the overall configuration of an engine according to an embodiment of the present invention, and FIG. 2 is a cross-sectional view showing the configuration of the engine body 1. The engine shown in these drawings is a multi-cylinder gasoline engine, and the engine body 1 is provided with a plurality of cylinders (four cylinders 1A to 1D in the illustrated example), and each cylinder 1A to 1D has a piston 2 respectively. (FIG. 2) is inserted. The piston 2 is connected to the crankshaft 3 via a connecting rod 4, and the crankshaft 3 rotates about the axis in accordance with the reciprocating motion of the piston 2.

  A combustion chamber 5 is formed above the piston 2, an intake port 6 and an exhaust port 7 are opened in the combustion chamber 5, and an intake valve 8 and an exhaust valve 9 that open and close the ports 6 and 7 are the upper part of the engine body 1. Is provided. The engine shown in the figure is a so-called double overhead camshaft (DOHC) engine, and two intake valves 8 and two exhaust valves 9 are provided for each cylinder. Further, cam shafts 40 and 41 (FIG. 2) that rotate in conjunction with the crankshaft 3 are provided above the intake valve 8 and the exhaust valve 9, and a plurality of cams attached to the camshafts 40 and 41, respectively. The intake / exhaust valves 8 and 9 are individually opened and closed by 40a and 41a.

  The engine body 1 is provided with a VVT 42 as a variable valve timing mechanism that makes it possible to change the closing timing of the intake valve 8.

  The VVT 42 includes, for example, a phase-type variable mechanism, and is configured to be able to change the rotational phase of the intake camshaft 40 with respect to the crankshaft 3 in accordance with the operating state of the engine. Since the structure of the VVT 42 has been known in the past, a detailed description thereof will be omitted. For example, the camshaft can be rotated relatively between the cam pulley and the camshaft, which transmit the rotation of the crankshaft via a timing belt. The phase changing member is incorporated, and this member is driven hydraulically or electrically.

  As the variable valve timing mechanism, a variable mechanism that changes the closing timing of the intake valve 8 by changing the valve lift amount may be provided. In addition, by using such a variable valve lift amount mechanism in combination with the above-described phase-type variable mechanism, the effective compression ratio change control and the overlap amount control of the intake / exhaust valves 8 and 9 can be simultaneously performed. It may be possible to do this.

  Here, unlike the general gasoline engine, the compression ratio of the engine of this embodiment is set to be considerably high. Specifically, the geometric compression ratio of a general gasoline engine is about 9 to 11, whereas the geometric compression ratio of the engine of this embodiment is set to about 18.

  As shown in FIGS. 1 and 2, the engine body 1 includes an injector 10 that directly injects fuel (gasoline) into the combustion chamber 5, and a spark plug 11 that discharges a spark for ignition into the combustion chamber 5. One is provided for each cylinder. In the illustrated example, an injector 10 is provided so as to face the combustion chamber 5 from the side of the intake side, and an ignition plug 11 is provided so as to face the combustion chamber 5 from above.

  The spark plug 11 is electrically connected to an ignition circuit device 12 that generates electric power for spark discharge, and a spark is generated from the spark plug 11 at a predetermined timing in response to power supply from the ignition circuit device 12. It is supposed to be discharged.

  As shown in FIG. 2, the engine body 1 is provided with an engine rotation speed sensor 61 that detects the rotation speed of the crankshaft 3 and a water temperature sensor 62 that detects the temperature of cooling water of the engine. Yes.

  An intake passage 13 and an exhaust passage 19 are connected to the intake port 6 and the exhaust port 7 of the engine body 1, respectively.

  The intake passage 13 is a passage for supplying combustion air to the combustion chamber 5. As shown in FIG. 1, the intake passage 13 is provided in common with a plurality of branch passage portions 14 branched for each cylinder and upstream thereof. And a common passage portion 15.

  The exhaust passage 19 is a passage for discharging burned gas (exhaust gas) generated in the combustion chamber 5, and similarly to the intake passage 13, a plurality of branch passage portions 20 branched for each cylinder, And a common passage portion 21 provided in common on the downstream side.

  An air flow sensor 60 for detecting the flow rate of intake air passing through the common passage portion 15 is provided in the common passage portion 15 of the intake passage 13 upstream of the compressor 27 described later.

  The common passage portion 15 is provided with a throttle valve 16 for adjusting the intake air amount. The throttle valve 16 is an electronically controlled throttle valve that is opened and closed by an actuator 17. That is, the opening degree of an accelerator pedal (not shown) to be depressed by the driver is detected by an accelerator opening degree sensor 63 (FIG. 3), and an ECU 50 (described later) is detected according to the detected opening degree, the engine operating state, and the like. 3) calculates an appropriate opening degree of the throttle valve 16, and a drive signal corresponding to the opening degree is inputted to the actuator 17 so that the throttle valve 16 is driven to open and close.

  The common passage portion 21 of the exhaust passage 19 is provided with a catalytic converter 24 incorporating a three-way catalyst, and harmful components in the exhaust gas passing through the exhaust passage 19 are purified by the action of the catalytic converter 24. It has become so.

  As shown in FIG. 1, the engine of this embodiment is provided with a supercharger 25 for pressurizing intake air.

  The supercharger 25 includes a turbine 26 provided in the common passage portion 21 of the exhaust passage 19, a compressor 27 provided in the common passage portion 15 of the intake passage 13, and a connection for connecting the turbine 26 and the compressor 27. It has a shaft 28 and an electric motor 29 that rotationally drives the connecting shaft 28. When the turbine 26 rotates in accordance with the energy of the exhaust gas, the compressor 27 rotates at a high speed in conjunction with this, whereby the air (intake air) passing through the intake passage 13 is pressurized and the combustion chamber 5 is pressurized. The electric motor 29 is driven as necessary to assist the rotation of the compressor 27.

  The compressor 27 comprises a relatively large impeller, and the supercharger 25 that pressurizes the intake air by such a large compressor 27 is particularly high in a high rotation or high load range where the energy of exhaust gas is large. Demonstrate performance. Further, when necessary, rotation assist by the electric motor 29 is performed, so that intake air is pressurized with excellent responsiveness.

  An intercooler 18 is provided on the downstream side of the compressor 27 in the common passage portion 15 of the intake passage 13 to cool the air whose temperature has increased due to supercharging.

  FIG. 3 is a block diagram showing an engine control system. The ECU 50 shown in the figure is a control device for comprehensively controlling each part of the engine, and includes a well-known CPU, ROM, RAM, and the like.

  The ECU 50 receives detection signals from various sensors. That is, the ECU 50 is electrically connected to the air flow sensor 60, the engine speed sensor 61, the water temperature sensor 62, and the accelerator opening sensor 63, and detection signals from these various sensors are sequentially input to the ECU 50. It has become so.

  The ECU 50 is also electrically connected to the injector 10, the ignition circuit device 12 for the spark plug 11, the actuator 17 for the throttle valve 16, the electric motor 29 for the supercharger 25, and the VVT 42. Each of the devices is configured to output a drive control signal.

  FIG. 4 is a diagram showing a control map that is referred to when the ECU 50 controls the engine. In this figure, the HCCI area A set in a relatively wide area on the low rotation side is an operation area where combustion by compression self-ignition is executed, and the SI area B set in the other high rotation area is This is an operating region where combustion by spark ignition is executed. That is, in the HCCI region A set in the low / medium speed rotation region of the engine, the air-fuel mixture (premixed gas) mainly based on the fuel injected during the intake stroke self-ignites before and after the compression top dead center. On the other hand, the air-fuel mixture is forcibly ignited by spark ignition from the spark plug 11 in the SI region B on the higher rotation side than this.

  Combustion by compression self-ignition performed in the HCCI region A is combustion in which self-ignition occurs at various locations in the combustion chamber 5 at the same time and has a shorter combustion period than combustion by spark ignition that has been conventionally performed. It is said that high thermal efficiency can be obtained. Further, combustion by compression self-ignition has an advantage that the combustion temperature is lower than that by spark ignition and the amount of NOx produced is greatly reduced.

  In particular, in this embodiment, as shown in FIG. 4, the HCCI region A is set over the entire load direction from no load to the maximum load, and if the rotational speed region is lower than the SI region B, Regardless of the value of the engine load (requested torque based on the accelerator opening), compression self-ignition combustion is performed uniformly. Normally, it is said that it is difficult to continue compression self-ignition combustion up to the high load range of the engine in terms of combustion controllability, but as will be described later, in this embodiment, the engine load is lean over the entire engine load range. While maintaining the air-fuel ratio, the compression self-ignition combustion can be continued up to the maximum load of the engine by, for example, retarding the fuel injection timing from the injector 10 in a high load region.

  On the other hand, in the SI region B set higher than the HCCI region A, the air-fuel mixture is forcibly ignited by spark ignition because the time from the compression stroke to the expansion stroke increases as the engine speed increases. This is because the time required for the self-ignition of the air-fuel mixture cannot be sufficiently secured, and it becomes difficult to reliably make the air-fuel mixture self-ignite. Therefore, in this embodiment, the engine is switched from compression self-ignition to combustion by spark ignition in the SI region B where the engine speed is relatively high.

  Returning to FIG. 3 again, specific functions of the ECU 50 will be described. The ECU 50 has valve timing control means 51, injector control means 52, ignition control means 53, and supercharging control means 54 as its main functional elements.

  The valve timing control means 51 controls the operation of the VVT 42 to appropriately change the closing timing of the intake valve 8 according to the operating state of the engine. That is, the intake valve 8 is normally closed near the retarded side of the intake bottom dead center (timing slightly past the intake bottom dead center), but depending on the operating state of the engine, the valve timing control means 51 When the VVT 42 is driven, the closing timing of the intake valve 8 is set much later than the intake bottom dead center. As a result, the substantial start timing of the compression stroke is delayed, and the compression ratio (effective compression ratio) of the engine is reduced accordingly. Therefore, in this embodiment, the means (variation ratio adjusting means) for variably setting the compression ratio of the engine includes the VVT 42 for changing the closing timing of the intake valve 8 and the valve timing for controlling the operation thereof. The control means 51 is comprised.

  The reduction control of the effective compression ratio by the compression ratio adjusting unit as described above is executed when the engine operating state shifts from the HCCI region A to the SI region B. That is, in this embodiment, in the HCCI region A where compression self-ignition combustion is performed, the effective compression ratio is uniformly set to 18 (that is, approximately the same value as the geometric compression ratio), while from the HCCI region A. On the other hand, in the high rotation side SI region B, in order to perform forced combustion by spark ignition while avoiding abnormal combustion such as knocking, control for reducing the effective compression ratio from 18 to about 15 is executed.

  The injector control means 52 controls the injection timing and injection amount (injection time) of the fuel injected from the injector 10 into the combustion chamber 5 by controlling the fuel injection operation by the injector 10. Then, a predetermined amount of fuel corresponding to the operating state is injected from the injector 10 under the control of the injector control means 52, and the injected fuel is mixed with the intake air, so that the combustion chamber 5 has a desired air-fuel ratio. An air-fuel mixture is generated.

  In particular, in this embodiment, when the operating state of the engine is in the HCCI region A where compression self-ignition combustion is performed, the fuel injection timing from the injector 10 is changed according to the engine load, or the fuel injection By performing the control to divide the number of times into a plurality of times, the injector control means 52 executes the self-ignition timing of the air-fuel mixture and the combustion period of the air-fuel mixture (the crank angle range from the start point to the end point of the combustion reaction). Is properly controlled. For example, if a large amount of fuel is injected all at once from the injector 10 on the high load side of the HCCI region A, the combustion becomes steep and the combustion period becomes excessively short, which may increase combustion noise. . Therefore, when the engine load increases to some extent in the HCCI region A, the injector control means 52 divides the fuel injection from the injector 10 into a plurality of times (more specifically, a part of the fuel is taken in the intake stroke). The combustion period due to compression self-ignition is moderately moderated so that the combustion period falls within a desired crank angle range. Further, the injector control means 52 intends the timing (ignition timing) at which self-ignition of the air-fuel mixture begins by retarding the timing (injection timing) at which fuel injection starts on the high load side of the HCCI region A. Therefore, it has a function of preventing the excessive increase in the in-cylinder pressure.

  The ignition control means 53 controls the timing (ignition timing) at which the spark plug 11 performs spark discharge by controlling the power supply from the ignition circuit device 12 to the spark plug 11.

  The supercharging control means 54 drives the electric motor 29 for the supercharger 25 as necessary, thereby controlling the supercharger 25 so as to obtain an appropriate supercharging characteristic according to the operating state. Is.

(2) Specific Example of Engine Combustion Control Next, combustion control in the HCCI region A where compression self-ignition combustion is performed will be described in detail. FIGS. 5A to 5E show the engine compression ratio (effective compression ratio) ε, the excess air ratio λ with respect to the stoichiometric air-fuel ratio, the ignition timing θig of the air-fuel mixture under the control of the ECU 50, It is a figure which shows how the combustion period (DELTA) (theta) of air-fuel | gaseous mixture and the supercharging amount QC by the supercharger 25 change according to an engine load (requested torque based on an accelerator opening degree). In these drawings, the load value X on the horizontal axis indicates the load threshold at which the ignition timing θig is changed, and Xm indicates the maximum load of the engine. Hereinafter, a range from no engine load to the load threshold value X (hereinafter referred to as a predetermined load X) is referred to as a load region A1, and a range from the predetermined load X to the maximum load Xm is referred to as a load region A2.

  First, as shown in FIG. 5A, the compression ratio ε is kept constant at ε = 18 over the entire engine load (both load regions A1 and A2).

  As shown in FIG. 5B, the excess air ratio λ with respect to the stoichiometric air-fuel ratio is set to fall within a predetermined range with λ = 3 as the upper limit over the entire engine load. For example, the excess air ratio λ gradually decreases from λ = 3 as the engine load changes from no load to an increasing direction. This is because while it is necessary to gradually increase the fuel injection amount in accordance with the load, the supercharging amount QC cannot be increased so as to be completely proportional thereto. However, as will be described later, when the load region A2 is equal to or greater than the predetermined load X, the supercharging amount QC is greatly increased. Therefore, the excess air ratio λ starts to increase in the load region A2, and returns to λ = 3. . As described above, the excess air ratio λ fluctuates within a certain range, but does not fall below λ = 2 at least, and in the illustrated example, it is maintained at a value relatively close to λ = 3 over the entire engine load. The

  As for the ignition timing θig of the air-fuel mixture, as shown in FIG. 5C, MBT (Minimum Advance for Best Torque) as an ignition timing at which the maximum torque can be obtained in the load range A1 from no load to the predetermined load X. On the other hand, when shifting to the load region A2 that is equal to or greater than the predetermined load X, the ignition timing θig is gradually retarded as the load increases, and is retarded by 3 ° with respect to the MBT at the maximum load Xm. Such retard of the ignition timing θig is performed by retarding the fuel injection timing from the injector 10 with respect to the injection timing corresponding to the MBT.

  That is, in order to start igniting the air-fuel mixture at the ignition timing MBT at which the maximum torque can be obtained, it is necessary to inject at least part of the fuel during the intake stroke and mix the fuel and air over a sufficient period of time. However, when it is desired to retard the ignition timing θig with respect to the MBT as described above, the injection start timing may be retarded so that fuel is mainly injected during the compression stroke. As a result, the time during which the fuel is exposed to high temperature and high pressure is shortened, the mixture is difficult to ignite, and the ignition timing θig is delayed from the MBT. The ignition timing MBT at which the maximum torque is obtained is theoretically specified as one crank angle, but the MBT in this specification is not intended to be strictly limited to the theoretically obtained MBT value. For example, it is assumed that the value has a predetermined width that allows an error of about ± 1 °.

  The combustion period Δθ of the air-fuel mixture is maintained at Δθ = 20 ° over the entire engine load, as shown in FIG.

  Finally, the supercharging amount QC by the supercharger 25 is gradually increased in accordance with the load as shown in FIG. 5 (e). In particular, the ignition timing θig is more than a predetermined load X at which the ignition timing θig starts to be retarded (load In the area A2), the supercharging amount QC is further increased.

B. Verification of Examples (1) Overall Guidelines According to the study by the present inventor, in the HCCI region A where compression self-ignition combustion is performed, the engine conforms to the control characteristics as shown in FIGS. 5 (a) to 5 (e). By controlling this, it is possible to prevent the in-cylinder pressure from rising excessively in a high load range while maintaining the thermal efficiency of the engine as high as possible. Below, the content of the research by this inventor who came to obtain such a conclusion is demonstrated.

  As shown in FIG. 6, in order to improve the thermal efficiency of the engine, it is necessary to reduce at least one of the four loss factors of exhaust loss, cooling loss, pump loss, and mechanical resistance. The inventor of the present application pays attention to the exhaust loss and the cooling loss, and as a control parameter related to these losses, the thermal efficiency can be further improved by controlling the compression ratio ε, the ignition timing θig, the combustion period Δθ, and the specific heat ratio. We considered whether it can be effectively improved. The specific heat ratio is closely related to the excess air ratio λ of the air-fuel mixture, and hence the excess air ratio λ will be taken as a parameter to be controlled instead of the specific heat ratio.

(2) Verification of thermal efficiency in the partial load region FIG. 8 shows the compression ratio ε, the excess air ratio λ with respect to the stoichiometric air-fuel ratio, and the mixture at the representative point R (see FIG. 7) set in the partial load region of the engine. It is a figure which shows the calculation result of the thermal efficiency (illustration thermal efficiency) at the time of changing each combustion period (DELTA) (theta). The calculation result here indicates the thermal efficiency when the engine is operated at the representative point R of the engine rotational speed Ne = 2000 rpm and the indicated mean effective pressure Pi = 300 kPa, as shown in FIG. The ignition timing θig is constant at MBT (ignition timing at which the torque becomes maximum).

  FIG. 8 shows the thermal efficiency in order from the left when the combustion period Δθ of the air-fuel mixture is set to 10 °, 20 °, 35 °, and 60 °. According to this figure, the maximum value of the indicated thermal efficiency is about 46% when Δθ = 60 ° and about 47% when Δθ = 35 °, whereas it is about Δθ = 10 ° and 20 °. About 48%. From this, it can be seen that the combustion period Δθ may be set to about 10 ° to 20 ° in order to further increase the thermal efficiency.

  However, when the thermal efficiency in the combustion period Δθ = 10 °, 20 ° is compared in detail, the range of the maximum value (48%) of the indicated thermal efficiency is slightly wider when Δθ = 10 °, and the overall distribution is You can see that there is no big difference between them. That is, even if the combustion period Δθ is shortened from 20 ° to 10 °, the effect of improving the thermal efficiency is hardly obtained. This is presumably because when the combustion period is shortened, the rate of increase of the in-cylinder pressure / temperature increases, thereby increasing the cooling loss. Moreover, it is assumed that it is difficult to actually control the combustion period Δθ to about 10 °. Thus, it can be said that Δθ = 20 ° is the target combustion period.

  When Δθ = 20 ° as described above, according to the corresponding graph (second graph from the left), the values of the compression ratio ε and the excess air ratio λ that give the maximum value (48%) of the indicated thermal efficiency are ε = 18 and λ = 5.

  However, according to the above graph, when the compression ratio ε = 18, the thermal efficiency is remarkably improved by changing the excess air ratio λ up to about λ = 3, and the range exceeding λ = 3. Thus, it can be seen that the thermal efficiency changes only slowly even when λ is changed. For example, when the excess air ratio λ is increased from λ = 2 → 3 at the compression ratio ε = 18, the thermal efficiency increases from 45% to 47% (that is, improves by 2%), but further air increases from λ = 3 → 5. Even if the excess rate is increased significantly, the thermal efficiency only rises to 48% (i.e. only improves by about 1%).

  Thus, when the excess air ratio exceeds λ = 3, the improvement in thermal efficiency is greatly reduced. Next, the reason for this will be considered. 9 and 10 are diagrams showing calculation results of exhaust loss and cooling loss that occur when the engine is operated under the same conditions as in FIG.

  First, as shown in FIG. 9, the engine exhaust loss is shorter when the combustion period Δθ is shorter, the compression ratio ε is higher, and the excess air ratio λ is larger. This is because when the combustion period Δθ is short and the compression ratio ε is large, the expansion period after the combustion is completed becomes longer and more work can be taken out, so that less energy is thrown away into the exhaust gas. This is because the larger the excess air ratio λ (that is, the leaner the air-fuel ratio), the lower the temperature of the exhaust gas, and the less energy is thrown away into the exhaust gas.

  That is, according to FIG. 9, it can be seen that the exhaust loss decreases as the excess air ratio λ increases for the same combustion period and the same compression ratio. From this, it can be considered that the thermal efficiency peaking phenomenon seen in FIG. 8 (decrease in the improvement of thermal efficiency when λ is over 3) is not caused by exhaust loss.

  On the other hand, the engine cooling loss increases as the combustion period Δθ is shorter and the compression ratio ε is higher, as shown in FIG. As for the excess air ratio λ, in the range of λ = 3 or less, the cooling loss decreases as λ increases. However, when λ = 3 is exceeded, the cooling loss starts to increase. Thus, when the excess air ratio λ becomes larger than λ = 3, the cooling loss is increased. Therefore, the thermal efficiency peak phenomenon shown in FIG. 8 occurs due to the cooling loss. It is thought that.

  Next, the reason why the cooling loss increases when the excess air ratio λ> 3 will be considered. Assuming that the cooling loss is Fc, Fc can be obtained by the following equation (1).

Fc = αS (T−Tw) (1)
Where α: heat transfer coefficient, S: combustion chamber surface area, T: gas temperature, Tw: combustion chamber wall temperature.

  In the above formula (1), the combustion chamber surface area S is always the same value if the model is the same engine, and the combustion chamber wall temperature Tw is always maintained at about 100 ° C. by the cooling water of the engine. Basically, there is no big change.

  On the other hand, the heat transfer coefficient α and the gas temperature T are values that vary depending on the combustion conditions. Among these, since the gas temperature T takes a lower value as the excess air ratio λ is larger, the cooling loss Fc should also be reduced in proportion thereto. From the above, it is considered that the increase in the cooling loss Fc with the excess air ratio λ> 3 is caused by the heat transfer coefficient α.

  Here, the heat transfer coefficient α can be obtained by the following equation (2).

α = 0.013D ^-0.2 P ^0.8 T ^-0.53 {2.28Up + c (P-Pm)} ^0.8 (2)
Here, D: cylinder bore, P: in-cylinder pressure, Up: average piston speed, c: combustion initial condition coefficient, Pm: motoring pressure.

  FIG. 11 shows the result of calculating the heat transfer coefficient α in the excess air ratio λ = 1, 3, 6 in relation to the crank angle based on the above equation (2). In addition, this FIG. 11 has shown the calculation result at the time of driving | running by rotational speed Ne = 2000rpm and the illustrated average effective pressure Pi = 300kPa (representative point R of FIG. 7) similarly to FIGS. Parameters other than the excess air ratio λ are compression ratio ε = 18, ignition timing θig = MBT, and combustion period Δθ = 20 °.

  As shown in FIG. 11, it can be seen that the heat transfer coefficient α increases as the excess air ratio λ increases. This is because when the excess air ratio λ increases and the air-fuel ratio becomes lean, the in-cylinder pressure P on the right side of the above equation (2) increases, and the heat transfer coefficient α increases in proportion thereto. Note that the heat transfer coefficient α increases as the in-cylinder pressure P increases. This is a temperature boundary layer (a layer of fluid whose temperature changes suddenly) formed along the wall surface of the combustion chamber. This is thought to be because heat transfer is promoted.

  FIG. 12 is a diagram showing the value of the integrated cooling loss value ΣdFc calculated based on the calculation result of the heat transfer coefficient α (FIG. 11) and the above equation (1). In FIG. 12, the larger the value of the integrated cooling loss value ΣdFc is on the negative side, the greater the cooling loss Fc is. According to this figure, the cooling loss Fc at the excess air ratio λ = 1, 3, 6 is the largest when λ = 1 and the smallest when λ = 3. The reason why the cooling loss Fc is large when λ = 1 is that combustion is performed at the stoichiometric air-fuel ratio and the combustion temperature is high.

  On the other hand, when the excess air ratio is increased from λ = 3 to λ = 6, the cooling loss Fc is increased. That is, if the excess air ratio λ is excessively increased, as described above, the in-cylinder pressure is excessively increased and the heat transfer coefficient α is significantly increased. This is considered to cause an increase in the cooling loss Fc.

  From the above, it can be said that it is not preferable to increase the excess air ratio beyond λ = 3 because the cooling loss Fc is increased. Of course, the larger the excess air ratio λ, the lower the exhaust gas temperature and the lower the exhaust loss (see FIG. 9), but after all, it is offset by the increase in the cooling loss Fc. As shown in FIG. 8, this is why the improvement of the thermal efficiency of the engine becomes sluggish when λ> 3.

  In addition, it is practically difficult to realize an extremely lean air-fuel ratio with an excess air ratio λ> 3 from the viewpoint of intake charge performance and the like. Therefore, it can be said that the excess air ratio λ should be set to λ = 3 in terms of practicality and thermal efficiency improvement.

  As for the parameters other than the excess air ratio λ, as described with reference to FIG. 8, the compression ratio ε = 18, the combustion period Δθ = 20 °, and the ignition timing θig = MBT should be targeted.

  Here, the presence or absence of practical problems caused by operating the engine under the above combustion conditions (ε = 18, λ = 3, θig = MBT, Δθ = 20 °) will be considered based on FIG. FIG. 13 shows the values of the in-cylinder pressure P, the rate of increase of the cylinder pressure P (pressure increase rate) dP / dθ, and the exhaust temperature Tex at the representative point R in FIG. It is a figure shown by the relationship with a corner. Among these, the exhaust temperature Tex is the exhaust gas temperature from the exhaust port 7 of the engine, and the temperature at the point E (that is, the temperature when the exhaust valve 9 is opened) in the illustrated TV diagram is shown. Equivalent to.

  In FIG. 13, the upper limit values of the in-cylinder pressure P, the pressure increase rate dP / dθ, and the exhaust temperature Tex are indicated by alternate long and short dash lines. Specifically, in FIG. 13, the upper limit value of the in-cylinder pressure P is set to a predetermined value in the range of 12 to 15 MPa, and the upper limit value of the pressure increase rate dp / dθ is set to a predetermined value in the range of 0.4 to 0.5 MPa / °. The upper limit value of the exhaust temperature Tex is set to about 1500K. This is in consideration of engine durability / reliability and combustion noise.

  According to FIG. 13, all of the values of the in-cylinder pressure P, the pressure increase rate dP / dθ, and the exhaust temperature Tex (the temperature at the point E) are within the upper limit value. It can be understood that the operation at ε = 18, λ = 3, θig = MBT, Δθ = 20 ° is sufficiently practical.

  As described above with reference to FIGS. 8 to 13, at the representative point R in the partial load region of the engine (rotational speed Ne = 2000 rpm, indicated mean effective pressure Pi = 300 kPa), the compression ratio ε = 18, the ignition timing θig. It was found that operating the engine under the conditions of = MBT, combustion period Δθ = 20 °, and excess air ratio λ = 3 is most desirable in terms of effectively improving thermal efficiency in consideration of practicality.

(3) Verification regarding load expansion Next, the combustion conditions at the representative point R, such as the compression ratio ε = 18, the excess air ratio λ = 3, the ignition timing θig = MBT, and the combustion period Δθ = 20 °, are continued to the high load range. Consider whether it is possible. FIG. 14 shows that when the engine operating state changes along the line L (constant rotation speed line) in FIG. 7, the maximum in-cylinder pressure Pmax, the maximum pressure increase rate dP / dθmax, and the exhaust gas temperature Tex are loaded ( It is a figure which shows how it changes according to the illustration average effective pressure Pi) of a horizontal axis. Hereinafter, these Pmax, dp / dθmax, and Tex may be collectively referred to as a combustion index value. Among these combustion index values, the maximum in-cylinder pressure Pmax is the maximum value of the in-cylinder pressure P shown in FIG. 13, and the maximum pressure increase rate dp / dθmax is the maximum value of the pressure increase rate dp / dθ. That is. In FIG. 14, the value of the indicated mean effective pressure Pi (about 1200 kPa) in the rightmost plot of each graph represents the maximum engine load (the value on the maximum load line M in FIG. 7). If the combustion index values (Pmax, dp / dθmax, Tex) do not exceed the upper limit in the range from no load to the maximum load, it can be determined that the engine can be operated without any problem.

  According to FIG. 14, it can be seen that, among the combustion index values, the maximum pressure increase rate dp / dθmax and the exhaust gas temperature Tex are within the upper limit values over the entire engine load, and are not particularly problematic in operation. On the other hand, the maximum in-cylinder pressure Pmax exceeds the upper limit value in the high load region of the engine. This is because if a lean air-fuel ratio of the excess air ratio λ = 3 is continued to the high load range, a considerable amount of air needs to be sent into the cylinder in the high load range, and an excessive increase in the cylinder pressure is unavoidable. it is conceivable that. From the above, it can be seen that the maximum in-cylinder pressure Pmax becomes a problem when the same combustion conditions as in the partial load region are continued up to the high load region, and it is necessary to change the combustion conditions to avoid this.

  Next, it will be examined how to change the combustion condition of the engine in order to keep the maximum in-cylinder pressure Pmax below the upper limit value. As described above, each graph of FIG. 14 shows the calculation results under the combustion conditions of compression ratio ε = 18, excess air ratio λ = 3, ignition timing θig = MBT, and combustion period Δθ = 20 °. Therefore, the combustion index values (Pmax, dp / dθmax, Tex) when the parameters of ε, λ, θig, and Δθ are changed are calculated, and it is examined whether they are within the ranges that do not exceed the upper limit values.

  FIG. 15 shows that the parameters ε, λ, θig, and Δθ are constant according to the load on the higher load side than the point at which the maximum in-cylinder pressure Pmax reaches the upper limit (that is, the indicated average effective pressure Pi≈1080 kPa). It is a figure which shows the change of Pmax, dp / d (theta) max, and Tex at that time changed by the ratio. In the figure, the broken line marked with ● represents the case where the compression ratio ε was lowered, the broken line marked with ▲ represents the case where the excess air ratio λ was reduced to enrich the air-fuel ratio, and the broken line marked with ▼ represents the ignition timing θig as retarded. In the case of (retarding), the broken line marked with ■ indicates the case where the combustion period Δθ is extended. According to these diagrams, when the indicated mean effective pressure Pi increased from about 1080 kPa to 1200 kPa, each parameter decreased the excess air ratio λ to ε = 18 → 15 when the compression ratio ε was decreased. Λ = 3 → 2.5, when ignition timing θig is retarded, θig = MBT → MBT−3 °, and when the combustion period Δθ is extended, Δθ = 20 ° → 40 °. . Also, the solid lines in the figure show the case where the above ε, λ, θig, and Δθ are kept constant without being changed.

  According to FIG. 15, the maximum in-cylinder pressure Pmax exceeds the upper limit value when the combustion period Δθ is extended, but otherwise falls within the upper limit value. The maximum pressure increase rate dp / dθmax exceeds the upper limit when the compression ratio ε is decreased and when the air-fuel ratio is enriched (that is, when the excess air ratio λ is decreased). In other cases, the value is below the upper limit. The exhaust temperature Tex remains below the upper limit value when any of ε, λ, θig, and Δθ is changed.

  FIG. 16 is a diagram showing changes in the indicated thermal efficiency when ε, λ, θig, and Δθ are changed as in FIG. 15. According to this figure, the thermal efficiency of the engine is highest when the ignition timing θig is retarded. Hereinafter, the combustion period Δθ is extended, the compression ratio ε is decreased, the air-fuel ratio is enriched (the excess air ratio λ is decreased). It can be seen that the thermal efficiency deteriorates in this order.

  FIG. 17 is a table summarizing the results of FIGS. 15 and 16. As shown in the figure, among the four options of reducing the compression ratio ε, reducing the excess air ratio λ (enriching), extending the combustion period Δθ, and retarding the ignition timing θig, reducing the compression ratio ε, excess air When either the reduction of the rate λ or the extension of the combustion period Δθ is selected, a problem arises in terms of the maximum in-cylinder pressure Pmax or the maximum pressure increase rate dp / dθmax. On the other hand, when the retard of the ignition timing θig is selected, no problem occurs in any of the maximum in-cylinder pressure Pmax, the maximum pressure increase rate dp / dθmax, and the exhaust temperature Tex, and the thermal efficiency ranks first. The best. From the above, it has been found that it is most advantageous to retard the ignition timing θig in the region where the indicated mean effective pressure Pi exceeds about 1080 kPa.

(4) Summary of verification results As described above, the following conclusions can be obtained from the contents of verification based on FIGS.

  (A) According to the calculation result of the thermal efficiency at the representative point R (FIG. 7) such as the engine rotational speed Ne = 2000 rpm and the indicated mean effective pressure Pi = 300 kPa, the compression ratio ε = 18, excess air in the partial load region of the engine It is most effective from the standpoint of practicality and thermal efficiency to operate under the combustion conditions of rate λ = 3, ignition timing θig = MBT, and combustion period Δθ = 20 °.

  (B) However, if the combustion conditions as described above (ε = 18, λ = 3, θig = MBT, Δθ = 20 °) are continued to the high load range of the engine, the maximum in-cylinder pressure Pmax becomes too large, and the engine Problems arise in terms of durability and reliability. In order to avoid this, it is necessary to retard the ignition timing θig in a high engine load range (Pi≈1080 kPa or more).

  (C) By retarding the ignition timing θig, all combustion index values (maximum in-cylinder pressure Pmax, the compression ratio ε, excess air ratio λ, combustion period Δθ) are compared with the case where the other parameters are changed. The thermal efficiency can be maintained at the highest value while suppressing the maximum pressure increase rate dp / dθmax and the exhaust temperature Tex) to the upper limit value or less.

  (D) FIG. 18 shows the values of the parameters ε, λ, θig, and Δθ when the above control is performed, and changes in the indicated thermal efficiency based on the values. By operating the engine under the conditions shown in this figure, it is possible to effectively avoid an excessive increase in the in-cylinder pressure in the high load region while maintaining the thermal efficiency as high as possible.

C. Summary and Effects of Examples As can be understood from the above description, the above-described A. The control of each parameter ε, λ, θig, Δθ, etc. in the embodiment of FIG. This is derived on the basis of the result obtained from the verification (FIG. 18). As a correspondence relationship between the two, the indicated mean effective pressure Pi = 1080 kPa at which the ignition timing θig starts to be retarded in FIG. 18 corresponds to the predetermined load X in FIG. In the following, A. above. The features and effects of the embodiment described in the above will be described together.

  In the above embodiment, as shown in FIG. 5, the air-fuel ratio of the air-fuel mixture is leaner than the stoichiometric air-fuel ratio (excess air ratio λ> 1) over the entire load direction of the HCCI region A where compression self-ignition combustion is performed. When the engine load increases to a predetermined load X or more in the HCCI region A, the self-ignition timing θig of the air-fuel mixture is retarded with respect to MBT, which is the ignition timing at which the torque becomes maximum. According to such a configuration, there is an advantage that compression self-ignition combustion under a lean air-fuel ratio can be performed in a wider operating region while preventing an excessive increase in the in-cylinder pressure.

  That is, in the above embodiment, by making the air-fuel ratio lean throughout the load direction of the HCCI region A, it is possible to effectively improve thermal efficiency and further improve fuel efficiency. However, if the lean air-fuel ratio is continued up to the high load side, it is necessary to increase the intake air amount significantly as the load increases. As a result, an excessive increase in the in-cylinder pressure P increases the durability and reliability of the engine. There is a risk of damage (see the graph of Pmax in FIG. 15). In order to deal with such a problem, in the above embodiment, the ignition timing θig of the air-fuel mixture is retarded with respect to the MBT in the load region (load region A2) equal to or higher than the predetermined load X, so that the in-cylinder pressure P is high as described above. It is possible to effectively prevent an excessive increase in the load range. As a result, the compression self-ignition combustion under the lean air-fuel ratio is continued to a higher load side (up to the maximum engine load Xm in the above embodiment). There is an advantage that the thermal efficiency can be greatly improved.

  In particular, when the excess air ratio λ with respect to the stoichiometric air-fuel ratio is set to λ≈3 in the entire load direction of the HCCI region A as in the above embodiment, as described with reference to FIGS. Both the engine exhaust loss and the cooling loss can be reduced in a balanced manner, and the thermal efficiency can be increased to a sufficient level within a feasible range. In addition, by maintaining a fairly lean air-fuel ratio of λ≈3 over the entire load direction, the combustion temperature can be sufficiently lowered to significantly reduce the amount of NOx discharged from the combustion chamber 5, and expensive NOx There is an advantage that the exhaust gas can be kept clean without providing a catalyst or the like.

  In the above embodiment, when the engine load increases to the predetermined load X in the HCCI region A, the ignition timing θig is retarded with respect to the MBT, while the compression ratio ε and the combustion period Δθ are equal to or greater than the predetermined load X. However, since the constant values (ε = 18, Δθ = 20 ° in the above embodiment) are maintained, parameters other than the ignition timing θig are maintained at appropriate values in terms of thermal efficiency. There is an advantage that an excessive increase in the in-cylinder pressure P can be effectively prevented while minimizing the decrease.

  Further, in the above embodiment, as shown in FIG. 5, in the HCCI region A, the supercharging amount QC by the supercharger 25 is increased according to the load, and in particular, a load equal to or higher than the predetermined load X at which the ignition timing θig is retarded. Range (load range A2), the supercharging amount QC is greatly increased, so that the engine output in the high load range is sufficiently secured by a large amount of supercharging, and the air-fuel ratio is made leaner to the high load range. There is an advantage that the thermal efficiency can be further improved continuously.

  In the above embodiment, the fuel injection timing from the injector 10 is retarded with respect to the injection timing corresponding to the MBT in order to retard the ignition timing θig with respect to the MBT at a predetermined load X or more. There is an advantage that an excessive increase in the in-cylinder pressure P can be effectively prevented by appropriately retarding the ignition timing θig in the high load region with a relatively simple configuration by simply changing the injection timing.

  Although not specifically described in the above embodiment, it is assumed that it is difficult to reliably cause the compression self-ignition under a significantly lean air-fuel ratio of the excess air ratio λ≈3, particularly in the low load region of the engine. Is done. Therefore, in such a case, ignition assistance may be performed using the spark plug 11. The ignition assist is to promote compression self-ignition by auxiliary spark discharge from the spark plug 11 before compression self-ignition starts. By performing such ignition assist, compression self-ignition combustion in the HCCI region A is stably performed, and misfire is reliably prevented.

  In addition, the ignition assist may be continuously performed up to a high load range of the engine. In such a case, the control for retarding the ignition timing θig of the air-fuel mixture with respect to the MBT above the predetermined load X is retarded with respect to the timing of the ignition assist (delayed from the ignition assist timing corresponding to the MBT). It may be realized with. Even in such a case, the control for retarding the ignition timing θig at a predetermined load X or more can be appropriately performed by changing the ignition assist timing, and it is effective that the in-cylinder pressure P is excessively increased in the high load region of the engine. There is an advantage that can be prevented.

  Further, in the above embodiment, in order to maintain a significantly lean air-fuel ratio of the excess air ratio λ≈3 over the entire engine load, it is necessary to send a considerably large amount of air into the cylinder, particularly in a high engine load range. For this reason, a high-performance supercharger equipped with a large compressor 27, an electric motor 29 for assisting rotation, and the like is used as the supercharger 25. However, the supercharging amount can be increased according to the operating state of the engine. In order to perform fine control, a plurality of superchargers having different supercharging characteristics may be provided in the engine, and these superchargers may be appropriately used according to the operating state of the engine.

  In the above embodiment, the excess air ratio λ with respect to the stoichiometric air-fuel ratio is maintained at λ≈3 over the entire engine load. However, for example, the amount of supercharging by the supercharger 25 is sufficiently increased due to problems such as cost. If it cannot be secured, the excess air ratio λ may be set to about λ = 2 in all or part of the load region. As shown in the second graph from the left in FIG. 8 (graph when Δθ = 20 °), even if the excess air ratio λ is reduced from about 3 to 2, for example, the thermal efficiency shown in the case of a compression ratio of 18 , The thermal efficiency can be sufficiently improved as compared with the conventional case.

  Moreover, in the said Example, as shown in FIG. 4, while setting HCCI area | region A in which compression self-ignition combustion is performed to the low / medium speed rotation area | region of an engine, SI area | region set to the high rotation side rather than this In B, combustion by spark ignition using the spark plug 11 is performed. However, for example, by using a fuel having an optimum property for compression self-ignition combustion, it is sufficient even in a high rotation region such as the SI region B. If it is possible to perform the compression self-ignition combustion controlled to the above, the operation by the compression self-ignition combustion may be performed in all the rotational speed ranges.

1A to 1D Cylinder 2 Piston 8 Intake valve 10 Injector 11 Spark plug 25 Supercharger 52 Injector control means X Predetermined load ε Compression ratio λ Air excess ratio θig Ignition timing Δθ Combustion period

Claims (7)

A method for controlling an engine in which compression self-ignition combustion is performed in which an air-fuel mixture is self-ignited by fuel injection from an injector and compression action of a piston in an HCCI region set in at least a part of an engine operation region,
When the air-fuel ratio of the air-fuel mixture is set to be leaner than the stoichiometric air-fuel ratio in the entire load direction of the HCCI region, and the engine load increases to a predetermined load or more in the HCCI region, the timing of the self-ignition of the air-fuel mixture is set. An engine control method characterized by retarding the MBT that is the ignition timing at which the torque becomes maximum. The engine control method according to claim 1,
An engine control method characterized by retarding the fuel injection timing from the injector more than the injection timing corresponding to the MBT when the engine load increases to the predetermined load or more in the HCCI region. The engine control method according to claim 1,
In order to assist the self-ignition of the air-fuel mixture, ignition assist for discharging a spark from the spark plug is performed in the HCCI region. An engine control method characterized by retarding the ignition assist timing corresponding to the MBT. The engine control method according to any one of claims 1 to 3,
A method for controlling an engine, characterized in that a combustion period of compression self-ignition combustion in the HCCI region is substantially constant regardless of whether or not the load is equal to or greater than the predetermined load. The engine control method according to any one of claims 1 to 4,
An engine control method, wherein an excess air ratio λ with respect to the stoichiometric air-fuel ratio is set to λ = 2 to 3 in the entire load direction of the HCCI region. In the engine control method according to any one of claims 1 to 5,
An engine control method, wherein supercharging is performed using a supercharger at least in a load range equal to or greater than the predetermined load. An apparatus for controlling an engine in which compression self-ignition combustion is performed in which an air-fuel mixture is self-ignited by fuel injection from an injector and compression action of a piston in an HCCI region set in at least a part of an engine operation region,
Injector control means for controlling the fuel injection operation from the injector,
In the entire load direction of the HCCI region, the air-fuel ratio of the air-fuel mixture is set to be leaner than the stoichiometric air-fuel ratio,
When the engine load increases above a predetermined load within the HCCI region, the injector control means executes control for retarding the self-ignition timing of the air-fuel mixture with respect to MBT, which is the ignition timing at which the torque becomes maximum. An engine control device.