JP2013029026A - Internal combustion engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide an internal combustion engine with a variable compression ratio mechanism, which prevents an operation state from becoming unstable when a mechanical compression ratio is changed.SOLUTION: The internal combustion engine includes the variable compression ratio mechanism that changes the mechanical compression ratio by change of the volume of a combustion chamber. An unburnt gas and a burned gas that has already burnt are included in a residual gas remaining in the combustion chamber without being exhausted at a previous combustion cycle rather than this combustion cycle. The internal combustion engine estimates the amount of the unburnt gas in this combustion cycle during the period of a transient operation in which the mechanical compression ratio has been changed, and sets an ignition timing in this combustion cycle based on the amount of the unburnt gas in this combustion cycle.

Description

  The present invention relates to an internal combustion engine.

  In the combustion chamber of the internal combustion engine, the air-fuel mixture is ignited in a compressed state. It is known that the compression ratio when compressing the air-fuel mixture affects the output torque and the fuel consumption. By increasing the compression ratio, the output torque can be increased or the fuel consumption can be reduced. On the other hand, it is known that if the compression ratio is too high, abnormal combustion such as knocking occurs. In the prior art, an internal combustion engine having a variable compression ratio mechanism capable of changing the compression ratio during an operation period is known. In addition to the compression ratio variable mechanism, there is known an internal combustion engine that includes a variable valve timing mechanism formed so that the opening / closing timing of the intake valve can be changed.

  In Japanese Patent Application Laid-Open No. 2010-90872, an exhaust gas recirculation mechanism that recirculates exhaust gas into the combustion chamber, and an exhaust gas that calculates an exhaust gas recirculation rate that is the ratio of the exhaust gas recirculated to the combustion chamber with respect to the total amount of gas supplied to the combustion chamber. An internal combustion engine that includes a recirculation rate calculating means and determines an ignition timing so that an increasing gradient of an amount of movement of the ignition timing toward the advance side with respect to an increase in the exhaust gas recirculation rate increases as the exhaust gas recirculation rate increases. Has been. Further, this publication discloses that when a mechanical compression ratio control mechanism for changing the mechanical compression ratio is provided, the ignition timing is determined to be a more advanced timing as the mechanical compression ratio is decreased.

JP 2010-90872 A

  The compression ratio variable mechanism can change the mechanical compression ratio by including a mechanism for changing the volume of the combustion chamber when the piston reaches the compression top dead center, for example. During the transient operation in which the mechanical compression ratio is changed, the volume of the combustion chamber gradually changes. For example, when the mechanical compression ratio is changed to be small, the volume of the combustion chamber when the piston reaches top dead center is increased.

  By the way, in the combustion chamber, when the air-fuel mixture is ignited, for example, a flame spreads around the spark plug. At this time, there is a region where the flame does not reach in the combustion chamber. When the air-fuel mixture burns in the combustion chamber, it remains as unburned gas in some areas without burning. The region where such combustion does not occur is called a quench zone. The size of the quench zone depends on the size of the combustion chamber. For this reason, if the mechanical compression ratio is changed by changing the size of the combustion chamber, the size of the quench zone changes, and as a result, the proportion of unburned gas also changes.

  On the other hand, in the exhaust stroke of the combustion cycle, not all of the gas burned in the expansion stroke is exhausted, and part of the burned gas remains until the next combustion cycle. This residual gas includes the above-mentioned unburned gas in addition to the already burned burned gas.

  During the transient operation period in which the mechanical compression ratio is changed, the ratio of the unburned gas contained in the residual gas changes due to the change in the size of the quench zone. For this reason, the operating state of the internal combustion engine may become unstable. For example, abnormal combustion such as knocking may occur during a period in which the mechanical compression ratio is changed.

  An object of the present invention is to provide an internal combustion engine that includes a compression ratio variable mechanism and suppresses an unstable operation state when the mechanical compression ratio is changed.

  The internal combustion engine of the present invention includes a variable compression ratio mechanism that changes the mechanical compression ratio by changing the volume of the combustion chamber. The residual gas remaining in the combustion chamber without being discharged in the combustion cycle prior to the current combustion cycle includes unburned gas and already burned burned gas. The internal combustion engine estimates the amount of unburned gas in the current combustion cycle during the transient operation while changing the mechanical compression ratio, and based on the amount of unburned gas in the current combustion cycle, Set the ignition timing in the cycle.

  In the above invention, the mechanical compression ratio in the current combustion cycle during the transient operation period is estimated, and the amount of fresh air flowing into the cylinder in the current combustion cycle and the mechanical compression ratio in the current combustion cycle are steady. Estimate the amount of unburned gas during operation, perform steady operation with the amount of unburned gas in the current combustion cycle during the transient operation, and the mechanical compression ratio in the current combustion cycle. The amount of fresh air for setting the ignition timing is calculated by correcting the amount of fresh air flowing into the cylinder in the current combustion cycle based on the amount of unburned gas when The ignition timing can be set based on the amount of air.

  In the above invention, the mechanical compression ratio in the current combustion cycle during the transient operation period is estimated, and based on the mechanical compression ratio and the amount of unburned gas in the combustion cycle prior to the current combustion cycle, The amount of unburned gas in the cycle can be estimated.

  In the above invention, the ignition timing in the steady operation in which the mechanical compression ratio is maintained constant is determined in advance, and during the transient operation period in which the mechanical compression ratio is reduced, the ignition timing is later than the ignition timing in the steady operation. Set the ignition timing on the corner side.

  According to the present invention, it is possible to provide an internal combustion engine that includes a variable compression ratio mechanism and suppresses an unstable operation state when the mechanical compression ratio is changed.

1 is a schematic view of an internal combustion engine in an embodiment. It is a general | schematic disassembled perspective view of the compression ratio variable mechanism in embodiment. In the internal combustion engine of the embodiment, it is a schematic cross-sectional view of a cylinder block and a crankcase portion when the mechanical compression ratio is a high compression ratio. In the internal combustion engine of the embodiment, it is a schematic cross-sectional view of a cylinder block and a crankcase portion when the mechanical compression ratio is a low compression ratio. In an internal combustion engine of an embodiment, it is an explanatory view of a ratio of gas contained in residual gas when changing a mechanical compression ratio at the time of steady operation. It is the schematic explaining the fresh air quantity for ignition timing setting in the operation control of embodiment. It is a flowchart of operation control in an embodiment. It is a graph explaining the relationship between the mechanical compression ratio in embodiment, and the ratio of burnt gas. 5 is a graph for estimating the in-cylinder temperature when the exhaust valve is closed to estimate the amount of residual gas in the operation control of the embodiment. It is a map of the ignition timing which makes the new air quantity in the operation control of embodiment a function.

  An internal combustion engine according to an embodiment will be described with reference to FIGS. In the present embodiment, an internal combustion engine disposed in a vehicle will be described as an example.

  FIG. 1 is a schematic view of an internal combustion engine in the present embodiment. The internal combustion engine in the present embodiment is a spark ignition type. The internal combustion engine includes an engine body 1. The engine body 1 includes a cylinder block 2 and a cylinder head 4. A piston 3 is disposed inside the cylinder block 2. The piston 3 reciprocates inside the cylinder block 2.

  The combustion chamber 5 is formed for each cylinder. An engine intake passage and an engine exhaust passage are connected to the combustion chamber 5. The engine intake passage is a passage for supplying fresh air to the combustion chamber 5. The fresh air is composed of air or a mixture of fuel and air. In the internal combustion engine in the present embodiment, since fuel is injected into the engine intake passage, fresh air flowing into the cylinder is composed of a mixture of fuel and air. The engine exhaust passage is a passage for discharging exhaust gas generated by the combustion of fuel from the combustion chamber 5.

  An intake port 7 and an exhaust port 9 are formed in the cylinder head 4. The intake valve 6 is disposed at the end of the intake port 7 and is configured to be able to open and close the engine intake passage communicating with the combustion chamber 5. The exhaust valve 8 is disposed at the end of the exhaust port 9 and is configured to be able to open and close the engine exhaust passage communicating with the combustion chamber 5. A spark plug 10 as an ignition device is fixed to the cylinder head 4. The spark plug 10 is formed to ignite fuel in the combustion chamber 5.

  The internal combustion engine in the present embodiment includes a fuel injection valve 11 for supplying fuel to the combustion chamber 5. The fuel injection valve 11 in the present embodiment is arranged so as to inject fuel into the intake port 7. The fuel injection valve 11 is not limited to this configuration, and may be arranged so that fuel can be supplied to the combustion chamber 5. For example, the fuel injection valve may be arranged to inject fuel directly into the combustion chamber.

  The fuel injection valve 11 is connected to the fuel tank 28 via an electronically controlled fuel pump 29 with variable discharge amount. The fuel stored in the fuel tank 28 is supplied to the fuel injection valve 11 by the fuel pump 29.

  The intake port 7 of each cylinder is connected to a surge tank 14 via a corresponding intake branch pipe 13. The surge tank 14 is connected to an air cleaner (not shown) through the intake duct 15. An air flow meter 16 that detects the amount of intake air is disposed inside the intake duct 15. A throttle valve 18 driven by a step motor 17 is disposed inside the intake duct 15. On the other hand, the exhaust port 9 of each cylinder is connected to a corresponding exhaust branch pipe 19. The exhaust branch pipe 19 is connected to the exhaust treatment device 21. The exhaust treatment device 21 in the present embodiment includes a three-way catalyst 20. The exhaust treatment device 21 is connected to the exhaust pipe 22.

  The internal combustion engine in the present embodiment includes an electronic control unit 31. The electronic control unit 31 in the present embodiment includes a digital computer. The electronic control unit 31 includes a RAM (random access memory) 33, a ROM (read only memory) 34, a CPU (microprocessor) 35, an input port 36 and an output port 37 which are connected to each other via a bidirectional bus 32. .

  The output signal of the air flow meter 16 is input to the input port 36 via the corresponding AD converter 38. A load sensor 41 is connected to the accelerator pedal 40. The load sensor 41 generates an output voltage proportional to the depression amount of the accelerator pedal 40. This output voltage is input to the input port 36 via the corresponding AD converter 38.

  The crank angle sensor 42 generates an output pulse each time the crankshaft rotates, for example, a predetermined angle, and this output pulse is input to the input port 36. The engine speed can be detected from the output of the crank angle sensor 42. Further, the crank angle can be detected from the output of the crank angle sensor 42. In the engine exhaust passage, a temperature sensor 43 as a temperature detector that detects the temperature of the exhaust treatment device 21 is disposed downstream of the exhaust treatment device 21. The output of the temperature sensor 43 is input to the input port 36 via the corresponding AD converter 38.

  The output port 37 of the electronic control unit 31 is connected to the fuel injection valve 11 and the spark plug 10 via the corresponding drive circuits 39. The electronic control unit 31 in the present embodiment is formed to perform fuel injection control and ignition control. That is, the fuel injection timing and the fuel injection amount are controlled by the electronic control unit 31. Further, the ignition timing of the spark plug 10 is controlled by the electronic control unit 31. The output port 37 is connected to a step motor 17 and a fuel pump 29 that drive the throttle valve 18 via a corresponding drive circuit 39. These devices are controlled by the electronic control unit 31.

  The intake valve 6 is formed to open and close as the intake cam 51 rotates. The exhaust valve 8 is formed to open and close as the exhaust cam 52 rotates. The internal combustion engine in the present embodiment includes a variable valve mechanism. The variable valve mechanism includes a variable valve timing device 53 that changes the opening / closing timing of the intake valve 6. The variable valve timing device 53 in the present embodiment is connected to the rotation shaft of the intake cam 51. The variable valve timing device 53 is controlled by the electronic control unit 31.

  The internal combustion engine in the present embodiment includes a variable compression ratio mechanism. The compression ratio of the internal combustion engine is determined depending on the volume of the combustion chamber when the piston reaches the compression top dead center. The variable compression ratio mechanism in the present embodiment is formed to change the compression ratio by changing the volume of the combustion chamber. The actual compression ratio, which is the actual compression ratio in the combustion chamber, is expressed by (actual compression ratio) = (combustion chamber volume + piston stroke volume when the intake valve is closed) / (combustion chamber volume).

  FIG. 2 is an exploded perspective view of the compression ratio variable mechanism of the internal combustion engine in the present embodiment. FIG. 3 is a first schematic cross-sectional view of the combustion chamber portion of the internal combustion engine. FIG. 3 is a schematic diagram when a high compression ratio is obtained by the variable compression ratio mechanism. In the internal combustion engine in the present embodiment, the lower structure including the crankcase and the cylinder block arranged on the upper side of the lower structure move relative to each other. The substructure in the present embodiment supports the cylinder block via a compression ratio variable mechanism. Further, the lower structure in the present embodiment supports the crankshaft.

  2 and 3, a plurality of protrusions 80 are formed below the side walls on both sides of the cylinder block 2. The protrusion 80 is formed with a cam insertion hole 81 having a circular cross section. A plurality of protrusions 82 are formed on the upper wall of the crankcase 79. The protrusion 82 is formed with a cam insertion hole 83 having a circular cross-sectional shape. The protrusion 82 of the crankcase 79 is fitted between the protrusions 80 of the cylinder block 2.

  The compression ratio variable mechanism in the present embodiment includes a pair of camshafts 84 and 85 as support shafts for the cylinder block. A circular cam 88 that is rotatably inserted into each cam insertion hole 83 is fixed to the cam shafts 84 and 85. The circular cam 88 is arranged coaxially with the rotation axis of each camshaft 84, 85. On the other hand, eccentric shafts 87 arranged eccentrically with respect to the rotation axis of the cam shafts 84 and 85 extend on both sides of each circular cam 88. On the eccentric shaft 87, another circular cam 86 is eccentrically attached to be rotatable. These circular cams 86 are arranged on both sides of the circular cam 88. The circular cam 86 is rotatably inserted into the corresponding cam insertion hole 81.

  The compression ratio variable mechanism includes a motor 89. Two worms 91 and 92 having spiral directions opposite to each other are attached to the rotating shaft 90 of the motor 89. Worm wheels 93 and 94 are fixed to the end portions of the camshafts 84 and 85, respectively. The worm wheels 93 and 94 are arranged so as to mesh with the worms 91 and 92. When the motor 89 rotates the rotating shaft 90, the camshafts 84 and 85 can be rotated in directions opposite to each other.

  Referring to FIG. 3, when the circular cams 88 arranged on the respective camshafts 84 and 85 are rotated in opposite directions as indicated by arrows 97, the eccentric shaft 87 faces the upper end of the circular cam 88. Moving. The circular cam 86 rotates in the opposite direction to the circular cam 88 as indicated by an arrow 96 in the cam insertion hole 81.

  FIG. 4 shows a second schematic cross-sectional view of the combustion chamber portion of the internal combustion engine in the present embodiment. FIG. 4 is a schematic diagram when a low compression ratio is achieved by the compression ratio variable mechanism. As shown in FIG. 4, when the eccentric shaft 87 moves to the upper end of the circular cam 88, the central axis of the circular cam 88 moves below the eccentric shaft 87. Referring to FIGS. 3 and 4, the relative position between crankcase 79 and cylinder block 2 is determined by the distance between the central axis of circular cam 86 and the central axis of circular cam 88. The cylinder block 2 moves away from the crankcase 79 as the distance between the central axis of the circular cam 86 and the central axis of the circular cam 88 increases. As the cylinder block 2 moves away from the crankcase 79 as indicated by an arrow 98, the volume of the combustion chamber 5 when the piston 3 reaches the compression top dead center increases.

  In the variable compression ratio mechanism in the present embodiment, the volume of the combustion chamber is variably formed by moving the cylinder block relative to the crankcase. In the present embodiment, a compression ratio determined only from the stroke volume of the piston from the bottom dead center to the top dead center and the volume of the combustion chamber is referred to as a mechanical compression ratio. In FIG. 3, the piston 3 has reached the compression top dead center, and the volume of the combustion chamber 5 is reduced. When the intake air amount is always constant, the compression ratio becomes high. This state is a state where the mechanical compression ratio is high. On the other hand, in FIG. 4, the piston 3 reaches the compression top dead center, and the volume of the combustion chamber 5 is increased. When the intake air amount is always constant, the compression ratio becomes low. This state is a state where the mechanical compression ratio is low. Thus, the internal combustion engine in the present embodiment can change the compression ratio during the operation period. For example, the compression ratio can be changed by a variable compression ratio mechanism according to the operating state of the internal combustion engine.

  Referring to FIG. 3, a relative position sensor 95 for detecting a relative position between crankcase 79 and cylinder block 2 is attached to a boundary portion between crankcase 79 and cylinder block 2. The relative position sensor 95 can detect the relative position between the crankcase 79 and the cylinder block 2 to detect the mechanical compression ratio. An output signal of the relative position sensor 95 is input to the electronic control unit 31.

  The variable compression ratio mechanism according to the present embodiment moves the cylinder block relative to the crankcase by rotating a circular cam having an eccentric rotation shaft. A variable compression ratio mechanism that can change the mechanical compression ratio by the mechanism can be employed.

  By the way, when the air-fuel mixture of fuel and air burns in the combustion chamber, there remains a portion that does not burn but becomes unburned gas. When the air-fuel mixture is ignited at the spark plug, a flame is generated, and the generated flame propagates toward the entire combustion chamber. At this time, a quench zone in which no flame reaches in the combustion chamber appears. For this reason, in addition to the burned gas in which the fuel is burned, the burned gas that has been burned includes unburned gas that remains without burning the fuel. Unburned gas includes fuel and air.

  In the internal combustion engine in the present embodiment, the piston rises to the top dead center in the exhaust stroke of the combustion cycle, but a part of the combustion gas does not flow out into the engine exhaust passage but remains in the combustion chamber. Residual gas remains until the next combustion cycle. For this reason, the gas sealed in the cylinder in each combustion cycle includes an air-fuel mixture of fuel and air and residual gas remaining in the past combustion cycle.

  The combustion gas includes burned gas and unburned gas, but when the mechanical compression ratio is changed, the size of the quench zone changes. For this reason, when the mechanical compression ratio changes, the ratio of burned gas and unburned gas contained in the residual gas changes.

  FIG. 5 is a schematic diagram illustrating the composition of residual gas remaining in the combustion chamber when the mechanical compression ratio changes. The piston 3 is supported on the crankshaft 59 via a connecting rod 58. As the crankshaft 59 rotates, the piston 3 reciprocates between the top dead center and the bottom dead center. FIG. 5 shows a case where the mechanical compression ratio ε is 20 as an example of the high compression ratio and a case where the mechanical compression ratio ε is 10 as the low compression ratio. Moreover, the example of the state of the steady operation in which each mechanical compression ratio is maintained constant is shown.

  In the variable compression ratio mechanism of the present embodiment, the cylinder block moves relative to the lower structure including the crankcase. For this reason, when the mechanical compression ratio changes, the volume of the combustion chamber changes. The gas remaining in the combustion chamber when the piston 3 reaches the top dead center in the exhaust stroke corresponds to the residual gas 64. The residual gas 64 includes unburned gas 64a in which fuel remains without being burned and burned burned gas 64b. Further, the stroke volume in which the piston 3 moves from the top dead center to the bottom dead center corresponds to the fresh air 63 newly supplied into the cylinder. The fresh air 63 and the unburned gas 64a include the fuel 61 and the air 62, thereby forming an unburned air-fuel mixture.

  The quench zone becomes smaller as the mechanical compression ratio changes from a high compression ratio to a low compression ratio. As the mechanical compression ratio becomes smaller, the proportion of the unburned gas 64a contained in the residual gas 64 becomes smaller. Comparing the ratio r of the burned gas in the residual gas 64 between the case of the high compression ratio and the case of the low compression ratio, the ratio r (ε = 20) of the burned gas 64b having the high compression ratio is the same as that of the low compression ratio. It can be seen that the ratio is smaller than the ratio r (ε = 10) of the fuel gas 64b.

  In the internal combustion engine in the present embodiment, the ignition timing in steady operation in which the mechanical compression ratio is maintained substantially constant is determined in advance. The ignition timing can be set based on the operating state of the internal combustion engine. The ignition timing in the present embodiment is set based on the amount of fresh air flowing into the cylinder and the engine speed. For example, as the amount of fresh air flowing into the cylinder increases, abnormal combustion such as knocking is more likely to occur, so that the ignition timing can be retarded.

  In the internal combustion engine in the present embodiment, in the steady operation in which the mechanical compression ratio is kept constant, the ignition timing is controlled without using the magnitude of the mechanical compression ratio as a function. The ignition timing in the steady operation in which the mechanical compression ratio is maintained constant is not limited to this form, and the mechanical compression ratio in the steady operation may be set as a function.

  The electronic control unit 31 stores in advance a map of ignition timing that is a function of the amount of fresh air flowing into the cylinder and the engine speed. In the present embodiment, the air flow meter 16 detects the amount of fresh air flowing into the cylinder, and the crank angle sensor 42 detects the engine speed. The ignition timing is set based on the detected amount of fresh air taken into the cylinder and the engine speed.

  In FIG. 6, the schematic explaining the operation control and comparative example of this Embodiment is shown. The internal combustion engine in the present embodiment performs control to retard the ignition timing from the ignition timing in the steady operation during the transient operation period in which the mechanical compression ratio is reduced. In the present embodiment, during the transient operation in which the mechanical compression ratio is changed, the amount of unburned gas in the current combustion cycle is estimated, and based on the amount of unburned gas in the current combustion cycle, Set the ignition timing for the current combustion cycle.

  Comparative Example 1 is an example when a steady operation is performed with the mechanical compression ratio ε maintained at 15. The ratio of the unburned gas 64a and the burned gas 64b in the residual gas 64 can be expressed as (1-r (ε = 15)): r (ε = 15).

  Comparative Example 2 is an example when the mechanical compression ratio ε decreases from 20 to 15. The mechanical compression ratio ε is the same as that in the comparative example 1, but in the comparative example 2, a transient operation in which the mechanical compression ratio is changed is performed. The amount of fresh air 63 flowing into the cylinder of Comparative Example 2 is the same as the amount of fresh air 63 flowing into the cylinder of Comparative Example 1. The total amount of residual gas 64 in Comparative Example 2 is the same as the total amount of residual gas 64 in Comparative Example 1.

  As described above, in the steady operation in which the mechanical compression ratio is kept constant, the ratio (1-r) of the unburned gas 64a in the residual gas 64 increases as the mechanical compression ratio increases. In the transient operation in which the mechanical compression ratio is lowered, it is influenced by the ratio of the high unburned gas 64a in the past combustion cycle. For this reason, the ratio of the unburned gas 64a during the transient operation of the comparative example 2 is larger than the ratio of the unburned gas 64a during the steady operation of the comparative example 1.

The ratio of the unburned gas 64a and the burned gas 64b in the residual gas 64 of Comparative Example 2 can be expressed as (1-r (ε _TDCfp )): r (ε _TDCfp ). The _burnt gas ratio r (ε _TDCfp ) indicates the burnt gas ratio when the piston reaches compression top dead center in the current combustion cycle during the period of transient operation.

  The amount of unburned gas 64a in Comparative Example 2 is greater than the amount of unburned gas 64a in Comparative Example 1. In Comparative Example 2, the amount of the unburned air-fuel mixture in the cylinder including the unburned gas 64a and the fresh air 63 is larger than that in Comparative Example 1. For this reason, abnormal combustion such as knocking may occur. Thus, even if the mechanical compression ratio is the same, the amount of unburned air-fuel mixture sealed in the combustion chamber during transient operation is larger than that during steady operation, and thus abnormal combustion may occur.

  FIG. 6 shows an example of operation control in the present embodiment. In the embodiment, the amount of fresh air 63a for setting the ignition timing is calculated by correcting the amount of fresh air 63 flowing into the cylinder. The amount of fresh air 63a for setting the ignition timing is larger than the amount of fresh air 63 of Comparative Example 1 during steady operation.

The amount of the fresh air 63a for setting the ignition timing is such that the sum of the fresh air 63 and the natural gas 64a during the transient operation of Comparative Example 2 is the sum of the corrected fresh air 63a and the unburned gas 64a in the embodiment. It is calculated to be the same. Further, the amount of the fresh air 63a for setting the ignition timing is such that the ratio of the unburned gas 64a and the fresh air 63a in the embodiment is the ratio of the unburned gas 64a and the fresh air 63 in the comparative example 1 during steady operation. It is calculated to be the same. In the case of the embodiment shown in FIG. 6, the ratio between the amount of unburned gas and the amount of fresh air is m _res (ε = 15) {1-r (ε = 15)}: m _new (ε = 15). Thus, the amount of fresh air 63a for setting the ignition timing is calculated. Here, the variable m _res is the amount of residual gas, and the variable m _new indicates the amount of fresh air flowing into the cylinder.

In the transient operation of Comparative Example 2, the amount of the unburned gas 64a in the current combustion cycle during the transient operation can be estimated by estimating the ratio of the unburned gas 64a and the burned gas 64b. . As the ratio r (ε _TDCfp ) of the _burnt gas 64b, the ratio of the burnt gas 64b corresponding to the mechanical compression ratio in the combustion cycle prior to the current combustion cycle can be employed. For example, the ratio of the burned gas 64b during steady operation corresponding to the mechanical compression ratio in the previous combustion cycle can be employed. The ratio of the unburned gas 64a can be estimated from the ratio of the burned gas 64b.

  In the internal combustion engine of the present embodiment, the ignition timing is set based on the amount of fresh air 63a for setting the ignition timing during the transient operation period in which the mechanical compression ratio is reduced. In the present embodiment, the ignition timing is set by a map stored in the electronic control unit 31 using the amount of the fresh air 63a for setting the ignition timing. Since the amount of the fresh air 63a for setting the ignition timing is larger than the amount of the fresh air 63 before correction, the ignition timing delayed from the ignition timing in the steady operation can be set.

  Thus, in the present embodiment, the amount of unburned gas 64a in the current combustion cycle during the transient operation of Comparative Example 2 and the mechanical compression ratio in the current combustion cycle of Comparative Example 1 are steady. Based on the amount of unburned gas 64a during operation, the amount of fresh air 63 flowing into the cylinder in the current combustion cycle is corrected, and the amount of fresh air 63a for setting the ignition timing is calculated. The ignition timing is set based on the calculated amount of fresh air 63a for setting the ignition timing.

  In the operation control of the internal combustion engine in the present embodiment, since the ignition timing is retarded when the mechanical compression ratio is lowered, the occurrence of abnormal combustion such as knocking can be suppressed. Further, the ignition timing can be retarded according to the amount of unburned gas contained in the residual gas. That is, the retard amount of the ignition timing can be increased as the amount of the unburned air-fuel mixture filled in the combustion chamber increases. Thus, the internal combustion engine in the present embodiment can suppress the operation state from becoming unstable when the mechanical compression ratio is changed.

  FIG. 7 shows a flowchart of operation control in the present embodiment. The flowchart shown in FIG. 7 can be repeated, for example, while the mechanical compression ratio is being changed.

  In step 101, it is determined whether or not the piston is located at the top dead center (compression top dead center) of the compression stroke. If the piston is not located at the top dead center of the compression stroke, this control is terminated. In step 101, when the piston is located at the top dead center of the compression stroke, the routine proceeds to step 102.

In step 102, the mechanical compression ratio ε _TDCfp1 when the piston is located at the top dead center of the compression stroke is detected. In this step, the mechanical compression ratio when the piston is located at the compression top dead center in the past combustion cycle is detected. In the present embodiment, the mechanical compression ratio when the piston is positioned at the compression top dead center in the previous combustion cycle is detected. In the present embodiment, the mechanical compression ratio ε can be detected by the relative position sensor 95 (see FIG. 3).

Next, in step 103, it is determined whether or not the piston is located at the bottom dead center of the intake stroke (or the bottom dead center of the compression stroke). In step 103, when the piston is not located at the bottom dead center of the intake stroke, this control is repeated. That is, it waits until the piston reaches the bottom dead center of the intake stroke (until the crank angle rotates 540 °). In step 103, if the piston has reached the bottom dead center of the intake stroke, the routine proceeds to step 104. In step 104, the mechanical compression ratio ε _BDCivc at this time is detected.

Next, in step 105, the mechanical compression ratio ε _TDCfp2 when the piston reaches the top dead center of the compression stroke in the current combustion cycle is estimated. The mechanical compression ratio ε _TDCfp2 when the piston reaches the top dead center of the compression stroke can be estimated using the mechanical compression ratio ε _TDCfp1 detected in step 102 and the mechanical compression ratio ε _BDCivc detected in step 104. In this combustion cycle, the mechanical compression ratio ε _TDCfp2 when the piston is located at the top dead center of the compression process can be estimated by the following equation (1).

Next, in step 106, the ratio r (ε _TDCfp1 ) of the burned gas in the residual gas when the piston is located at the top dead center of the compression stroke in the previous combustion cycle, and the piston in the current combustion cycle The ratio r (ε _TDCfp2 ) of the burned gas of the residual gas when located at the top dead center of the compression stroke is estimated.

  FIG. 8 shows a graph for explaining the ratio of burnt gas contained in the residual gas with respect to the mechanical compression ratio during steady operation. As the mechanical compression ratio ε increases, the ratio r of the burned gas in the residual gas decreases. The relationship between the mechanical compression ratio and the ratio of burned gas shown in FIG. 8 can be stored as a map in advance in an electronic control unit or the like.

Referring to FIG. 7, in step 106, the ratio r (ε _TDCfp1) of the burnt gas that corresponds to the mechanical compression ratio epsilon _TDCfp1 detected at step 102, corresponding to the mechanical compression ratio epsilon _TDCfp2 estimated in step 105 The ratio r (ε _TDCfp2 ) of burned gas can be read from the map.

Next, in step 107, the residual gas amount m _res (ε _TDCfp2 ) when the piston reaches the top dead center of the compression stroke in the current combustion cycle is estimated. The residual gas amount m _res depending on the magnitude of the mechanical compression ratio can be calculated by the following equation (2), for example.

Here, the variable P _EVC is the pressure in the cylinder when the exhaust valve is closed. The variable _VEVC is the volume in the cylinder when the exhaust valve is closed, and varies depending on the respective mechanical compression ratios. The constant R is a gas constant, and the variable _TEVC is the temperature in the cylinder when the exhaust valve is closed.

The variable P _EVC can be detected by arranging an in-cylinder pressure sensor for detecting the pressure in the cylinder. Or it can estimate using the output of a crank angle sensor, etc. The variable _VEVC can be calculated from the following equation (3), for example.

Here, the variable V _cyl indicates the displacement, the variable n _cyl indicates the number of cylinders, the variable EVC is the closing timing of the exhaust valve, and the unit is [rad ATDC].

In addition, the variable _TEVC can store, for example, a map in which the operating state of the internal combustion engine is a function in advance in an electronic control unit or the like. The variable _TEVC can be estimated by detecting the operating state of the internal combustion engine and reading it from the stored map. The variable T _EVC can be expressed by the following equation (4).

  Here, the variable NE is the engine speed, the variable KL is the charging efficiency, and the variable SA is the ignition timing.

FIG. 9 shows a graph for explaining a map for estimating the temperature _TEVC in the cylinder when the exhaust valve is closed. The temperature T _EVC is, for example, monotonically increasing with respect to the engine speed NE and monotonically increasing with respect to the charging efficiency KL. On the other hand, the temperature T _EVC decreases monotonously with respect to the ignition timing SA. The temperature T _EVC can be estimated from the map having such a relationship. As described above, the residual gas amount m _res at each mechanical compression ratio can be calculated from the above equation (2).

Referring to FIG. 7, next, in step 108, a fresh air amount m _new flowing into the cylinder in the current combustion cycle is detected. In the present embodiment, the amount of new air m _new flowing into the cylinder can be detected by, for example, the output of an air flow meter. The detection of the amount of fresh air flowing into the cylinder is not limited to this form, and any control can be employed.

Next, in step 109, the estimated new air amount _mmew is corrected to calculate the new air amount _{mnew, map} for setting the ignition timing. The new air amount m _{new, map} for setting the ignition timing can be calculated by the following equation (5).

Here, referring to FIG. 6, the variable [m _new + m _res (ε _TDCfp2 ) · {1−r (ε _TDCfp1 )}] in the equation (5) corresponds to the comparative example 2, and during the transient operation period This corresponds to the sum of fresh air 63 and unburned gas 64a in the current combustion cycle. As the ratio r of burnt gas in the current combustion cycle, the ratio r of burnt gas when the steady operation is performed at the mechanical compression ratio of the past combustion cycle is adopted. The variable [m _new + m _res (ε _TDCfp2 ) · {1-r (ε _TDCfp2 )}] in equation (5) corresponds to Comparative Example 1, and the steady operation is performed at the mechanical compression ratio in the current combustion cycle. Corresponds to the sum of the fresh air 63 and the unburned gas 64a. Thus, by multiplying the detected actual fresh air amount m _new by the ratio of the unburned mixture amount in the transient operation to the unburned mixture amount in the steady operation, the ignition timing setting The fresh air amount mnew _{, map} can be calculated.

  Referring to FIG. 7, next, at step 110, the engine speed is detected. In the present embodiment, the engine speed is detected by the output of the crank angle sensor 42 attached to the crankshaft.

Next, in step 111, the ignition timing in the current combustion cycle is set. In the present embodiment, the electronic control unit 31 stores a map of the ignition timing that uses the new air amount m _{new, map} for setting the ignition timing as a function.

FIG. 10 shows a map of the ignition timing using the new air amount for setting the ignition timing in this embodiment as a function. When the engine speed N and the new air amount m _{new, map} for setting the ignition timing are determined, the ignition timing FPA _mn can be set. In the present embodiment, in addition to the map of the ignition timing at the time of normal operation, a map that functions as a function of the new air amount for setting the ignition timing at the time of transient operation is stored in the electronic control unit. The map is not limited, and the ignition timing map during normal operation and transient operation may be shared.

  As described above, in the internal combustion engine in the present embodiment, at the time of transient operation in which the mechanical compression ratio is changed, it is possible to set an ignition timing obtained by correcting the ignition timing at the time of normal operation. By performing ignition at the corrected ignition timing in the current combustion cycle, abnormal combustion such as knocking can be suppressed.

  In the present embodiment, the amount of unburned gas in the current combustion cycle during the transient operation is estimated based on the mechanical compression ratio in the previous combustion cycle and the amount of unburned gas corresponding to the mechanical compression ratio. However, the present invention is not limited to this configuration, and the mechanical compression ratio and the amount of unburned gas in the combustion cycle prior to the current combustion cycle can be employed. In addition, the amount of unburned gas in the current combustion cycle during the transient operation can be estimated by arbitrary control.

  Further, in the present embodiment, the amount of unburned gas in the current combustion cycle during the transient operation period and the amount of unburned gas when performing steady operation at the mechanical compression ratio in the current combustion cycle. Based on the amount, the amount of fresh air flowing into the cylinder in the current combustion cycle is corrected to calculate the amount of fresh air for setting the ignition timing. Thus, it is possible to calculate the amount of fresh air for setting the ignition timing by correcting the amount of fresh air flowing into the cylinder in the current combustion cycle. For example, a correction coefficient having a function of the change amount of the mechanical compression ratio is determined in advance, and the amount of unburned gas when the steady operation is performed at the mechanical compression ratio in the current combustion cycle is estimated. By multiplying the amount of fuel gas by a correction coefficient, the amount of unburned gas after correction may be estimated, and the amount of fresh air for setting the ignition timing may be calculated based on the amount of unburned gas after correction. .

  Further, in the present embodiment, the ignition timing is set based on the amount of fresh air for setting the ignition timing. However, the present invention is not limited to this mode, and it is not necessary to calculate the amount of fresh air for setting the ignition timing. It doesn't matter. For example, the amount of unburned gas in the current combustion cycle may be estimated, and the ignition timing may be set directly using a map or the like based on the estimated amount of unburned gas.

  In the present embodiment, the transient operation in which the mechanical compressor is reduced has been described as an example. However, the present invention is not limited to this embodiment, and the present invention is also applied to the setting of the ignition timing in the transient operation in which the mechanical compression ratio increases. Can do. During the transient operation period in which the mechanical compression ratio is increased, the ignition timing can be set to an advance side with respect to the ignition timing in the steady operation.

  Further, the control of the ignition timing of the internal combustion engine in the present embodiment can be applied to control for setting the ignition timing when the volume of the combustion chamber changes when the piston reaches top dead center. For example, when sludge or the like gradually accumulates inside the combustion chamber as the internal combustion engine is used, the volume of the combustion chamber gradually decreases. In such an internal combustion engine, the volume of the combustion chamber when the piston reaches top dead center may be estimated, and the ignition timing may be changed based on the estimated volume of the combustion chamber. As an example in which the volume of the combustion chamber changes when the piston reaches top dead center, the temperature of the engine body can be exemplified in addition to the accumulation of sludge. The size of the combustion chamber may increase as the temperature of the engine body increases. In such an internal combustion engine, the temperature of the engine body is detected, the volume of the combustion chamber is estimated based on the temperature of the engine body, and the ignition timing may be changed based on the estimated volume of the combustion chamber. .

  In the above-described operation control, the order of the steps can be appropriately changed within a range in which the function of each step is not changed.

  In the present embodiment, the internal combustion engine disposed in the vehicle has been described as an example. However, the present invention is not limited to this embodiment, and the present invention can be applied to any internal combustion engine including a variable compression ratio mechanism.

  In the respective drawings described above, the same or corresponding parts are denoted by the same reference numerals. In addition, said embodiment is an illustration and does not limit invention. In the embodiment, the change shown in a claim is included.

DESCRIPTION OF SYMBOLS 1 Engine body 3 Piston 5 Combustion chamber 10 Spark plug 31 Electronic control unit 42 Crank angle sensor 61 Fuel 62 Air 63 Fresh air 63a Fresh air 64 Residual gas 64a Unburned gas 64b Burned gas 84, 85 Camshaft 86 Circular cam 87 Eccentricity Shaft 88 Circular cam 89 Motor 95 Relative position sensor

Claims (4)

It has a compression ratio variable mechanism that changes the mechanical compression ratio by changing the volume of the combustion chamber,
Residual gas remaining in the combustion chamber without being discharged in the combustion cycle prior to the current combustion cycle includes unburned gas and burned burned gas,
During the transient operation with the mechanical compression ratio changed, the amount of unburned gas in the current combustion cycle is estimated, and the ignition timing in the current combustion cycle is based on the amount of unburned gas in the current combustion cycle. An internal combustion engine characterized by setting. Estimate the mechanical compression ratio in this combustion cycle during the transient operation,
Estimate the amount of fresh air flowing into the cylinder during the current combustion cycle and the amount of unburned gas during steady operation at the mechanical compression ratio during the current combustion cycle,
Based on the amount of unburned gas in the current combustion cycle during the transient operation and the amount of unburned gas in steady operation at the mechanical compression ratio in the current combustion cycle, The amount of fresh air for setting the ignition timing is calculated by correcting the amount of fresh air flowing into the cylinder at
2. The internal combustion engine according to claim 1, wherein the ignition timing is set based on an amount of fresh air for setting the ignition timing. Estimate the mechanical compression ratio in this combustion cycle during the transient operation,
3. The amount of unburned gas in the current combustion cycle is estimated based on the mechanical compression ratio and the amount of unburned gas in the combustion cycle prior to the current combustion cycle. Internal combustion engine. The ignition timing in steady operation where the mechanical compression ratio is kept constant is predetermined,
4. The ignition timing that is retarded from the ignition timing in steady operation is set during a transient operation in which the mechanical compression ratio is reduced. 5. Internal combustion engine.

Images