JP6464962B2 - Engine control device - Google Patents

Description

  The present invention relates to an engine control apparatus, and more particularly to an engine control apparatus having a supercharger and a variable compression ratio mechanism.

  Some control the compression ratio of the variable compression ratio mechanism so that a high target compression ratio is obtained on the low load side and low on the high load side (see Patent Document 1).

JP 2006-46194 A

  By the way, for engines that employ a compression ratio variable mechanism in addition to a supercharger, it is an issue to further improve the response speed of the supercharging pressure in the supercharging region from the viewpoint of power performance and response performance. In response to this problem, the present inventor has improved the response speed of the supercharging pressure while maintaining the response speed of the compression ratio of the variable compression ratio mechanism. New problems have been identified.

  That is, at the time of deceleration and re-acceleration from a constant speed running state, the target compression ratio is switched from the low side to the high side by deceleration, and the target compression ratio is switched from the high side to the low side by re-acceleration immediately after that. . The compression ratio (actual compression ratio) of the variable compression ratio mechanism follows the target compression ratio thus changing. At this time, the time until the actual compression ratio reaches the target compression ratio after re-acceleration does not change before and after the boost pressure response speed is improved. On the other hand, by improving the response speed of the supercharging pressure, the time until the actual supercharging pressure reaches the target supercharging pressure after reacceleration becomes shorter than before the response speed of the supercharging pressure is improved. As a result, the timing at which the actual compression ratio reaches the target compression ratio after reacceleration is delayed from the timing at which the actual supercharging pressure reaches the target boost pressure after reacceleration. That is, at the timing when the actual boost pressure reaches the target boost pressure, a state in which the actual compression ratio is higher than the target compression ratio after re-acceleration appears. A state of high supercharging and a high compression ratio occurs, and knocking occurs. However, if the ignition timing is retarded or the actual boost pressure is limited to reach the target boost pressure after the reacceleration to avoid knocking, the deceleration reacceleration during constant speed driving Sometimes the engine response is significantly worse.

  Therefore, an object of the present invention is to provide a device that can suppress the occurrence of knocking even during deceleration and reacceleration during constant speed traveling.

The engine control apparatus of the present invention includes a supercharger, a compression ratio variable mechanism, a rotation speed detection means, a load detection means, a target compression ratio setting means, and a control means. The compression ratio variable mechanism can change the compression ratio of the engine. The rotational speed detecting means detects the rotational speed of the engine. The load detecting means detects an engine load. The target compression ratio setting means sets the target compression ratio on the high rotation speed side to be lower than the low rotation speed side even at the same load, based on the detected engine rotation speed and the detected engine load. The control means controls the compression ratio of the variable compression ratio mechanism so that the set target compression ratio is obtained. In addition, the target compression ratio setting means is configured to determine the time until the actual supercharging pressure reaches the set supercharging pressure or the actual engine during deceleration and reacceleration from a state where the vehicle is running at a constant speed in the high rotational speed region. The target in the high rotation speed region should not exceed the time until the torque reaches the set torque so that the time until the compression ratio of the variable compression ratio mechanism reaches the target compression ratio after re-acceleration is not exceeded. The compression ratio is set, and the compression ratio control means sets the target compression ratio set within the time until the actual boost pressure reaches the set boost pressure or the time until the actual engine torque reaches the set torque. To control the compression ratio of the variable compression ratio mechanism.

  As a comparative example, the actual response speed of the supercharging pressure is not improved in the supercharging region as described above, and in the present invention, particularly during deceleration reacceleration during constant speed traveling, before and after the deceleration reacceleration. The target compression ratio is set lower than in the comparative example. Since the target compression ratio is set low before and after the deceleration reacceleration, the amount of change in the target compression ratio before and after the deceleration reacceleration is smaller than the amount of change in the target compression ratio before and after the deceleration reacceleration in the comparative example. Become. If the amount of change in the target compression ratio before and after deceleration and re-acceleration is small, the time during which the operating point moves through the amount of change in the target compression ratio is shortened even if the response speed of the actual compression ratio is the same. Therefore, the timing at which the actual compression ratio reaches the target compression ratio after re-acceleration at the time of deceleration re-acceleration during constant speed traveling and the timing at which the actual boost pressure reaches the target boost pressure after re-acceleration are obtained. Can be equivalent. Thus, the actual compression ratio does not reach the target compression ratio after reacceleration at the timing when the actual supercharging pressure reaches the target boost pressure after reacceleration. Since a state of high supercharging and a high compression ratio does not occur, occurrence of knocking can be suppressed. If the occurrence of knocking can be suppressed, there is no need to retard the ignition timing or limit the actual supercharging pressure, so the torque or response in the supercharging region is improved during deceleration and reacceleration during constant speed driving. be able to.

It is a schematic block diagram of a reciprocating engine having a compression ratio variable mechanism. It is the schematic which shows the structure around a combustion chamber. It is a posture figure of each link in a high compression ratio position and a low compression ratio position. 1 is a schematic configuration diagram of an engine control system having a turbocharger and a compression ratio variable mechanism according to a first embodiment of the present invention. It is a characteristic figure of a target supercharging pressure. It is a characteristic view of the target compression ratio of the comparative example with respect to 1st Embodiment. It is a characteristic figure of the target compression ratio of a 1st embodiment. It is a characteristic view for demonstrating the 1st setting method of the target compression ratio of 1st Embodiment. It is a characteristic view for demonstrating the 2nd setting method of the target compression ratio of 1st Embodiment. It is a flowchart for demonstrating the setting of the target compression ratio of 1st Embodiment. It is a characteristic figure of the target compression ratio of a 2nd embodiment. It is a characteristic figure of the target compression ratio of the comparative example with respect to 3rd Embodiment. It is a characteristic figure of the target compression ratio of a 3rd embodiment.

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

(First embodiment)
FIG. 4 is a schematic configuration diagram of an engine control system having the turbocharger and the compression ratio variable mechanism according to the first embodiment of the present invention.

  Here, the reciprocating engine having the variable compression ratio mechanism will be described first. FIGS. 1 and 2 are schematic configuration diagrams of the reciprocating engine 1 having the variable compression ratio mechanism. Here, the compression ratio variable mechanism is mainly shown in FIG. 1, and the structure around the combustion chamber 21 of the engine 1 is mainly shown in FIG.

  The variable compression ratio mechanism shown in FIG. 1 is a multi-link type. Specifically, it is a mechanism for changing the compression ratio by changing the piston stroke. The engine 1 having a variable compression ratio mechanism has been previously proposed by the present applicant, and is known, for example, from Japanese Patent Application Laid-Open No. 2001-227367. Therefore, only the outline of the variable compression ratio mechanism will be described with reference to FIG.

  The crankshaft 3 is provided with a crank journal 4 that is rotatably supported by a main bearing (not shown) in a cylinder block 2 constituting a part of the engine body for each cylinder. Each crank journal 4 has an axis O that coincides with the axis (rotation center) of the crankshaft 3 and constitutes a rotating shaft portion of the crankshaft 3.

  In addition, the crankshaft 3 is eccentric from the axis O, is provided with a crankpin 5a provided for each cylinder, a crank arm 5b that connects the crankpin 5a to the crank journal 4, and opposite to the crankpin 5 with respect to the axis O. And a counterweight 5c that mainly reduces the primary rotational vibration component of the piston motion. The crank arm 5b and the counterweight 5c are integrally formed in this embodiment.

  In the present embodiment, the piston 10 slidably fitted to the cylinder 11 formed for each cylinder and the crank pin 5a described above are a plurality of link members, that is, the upper link 7 (first link). It is mechanically linked by the lower link 6 (second link). The upper end side of the upper link 7 is externally fitted to a piston pin 9 (first pin) fixedly provided on the piston 10 so as to be relatively rotatable around the axis Oc. Further, the lower link side of the upper link 7 and the one main body 6a of the lower link 6 that is substantially bisected can be relatively rotated around the axis Od by a connecting pin 8 (second pin) that passes through both of them. It is connected to.

  The lower link 6 is configured by attaching two main bodies 6a and 6b so as to sandwich the crank pin 5a, and is mounted so as to be relatively rotatable around the crank pin 5a and the axis Oe at the sandwiched portion. Yes. The other lower link body 6b, which is substantially divided into two parts, and the upper end side of the control link 12 (third link) can be relatively rotated around the axis Of by a connecting pin 13 (third pin) that passes through both of them. It is connected to.

  The lower end side of the control link 12 is swingable around its axis Ob (a fulcrum provided on the cylinder block) on a control shaft 14 having an eccentric cam portion 15 that is rotatably supported by the cylinder block 2. It is externally fitted and supported. That is, an eccentric cam portion 15 is rotatably provided on the outer periphery of the control shaft 14, and the axis Oa of the eccentric cam portion 15 is eccentric by a predetermined amount with respect to the axis Ob of the control shaft 14. The eccentric cam portion 15 is rotationally controlled by a compression ratio control actuator 17 via a worm gear 16 according to the operating state of the engine and is held at an arbitrary rotational position.

  With such a configuration, as the crankshaft 3 rotates, the piston 10 moves up and down in the cylinder 11 via the crankpin 5a, the lower link 6, the upper link 7 and the piston pin 9, and is connected to the lower link 6. The control link 12 swings with the swing axis Ob on the lower end side as a fulcrum.

  Further, the eccentric cam portion 15 is rotationally controlled by the compression ratio control actuator 17, so that the axis Ob of the control shaft 14 serving as the swing axis of the control link 12 is moved around the axis Oa of the eccentric cam portion 15. Rotation, that is, the swing center position Ob of the control link 12 moves with respect to the engine body (and crankshaft rotation center O). As a result, the stroke of the piston 10 changes, and the compression ratio of each cylinder of the engine is variably controlled. For reference, FIG. 3 schematically shows the postures of the three links 7, 6, and 12 at the piston top dead center position. The left side of FIG. 3 is the high compression ratio position, and the right side of FIG. 3 is the low compression ratio position. Each link posture.

  The greatest feature of this compression ratio variable mechanism is that the top dead center position (combustion chamber volume) of the piston 10 can be changed by the angular position control of the control shaft 14 (control shaft), and the function as a so-called variable compression ratio mechanism. To demonstrate. To explain this, the above-mentioned “compression ratio” is a value defined by the following equation.

Compression ratio = (combustion chamber volume at bottom dead center position) / (combustion chamber volume at top dead center position)
... (1)
As described with reference to FIGS. 1 and 3, in the compression ratio variable mechanism, the top dead center position of the piston 10 can be changed. Therefore, according to the compression ratio variable mechanism, the “compression ratio” in the expression (1) can be changed. It becomes.

  In addition, since the piston stroke characteristics can be made close to simple vibrations, the acceleration at the top and bottom dead center is substantially the same, and there is an effect of reducing vibrations that eliminates the need for a balancer shaft (four cylinders). Alternatively, the piston stroke characteristic can be set such that the piston acceleration on the top dead center side is smaller than the piston acceleration on the bottom dead center side. Such piston acceleration characteristics can be obtained with a multi-link mechanism composed of a plurality of link members as described above, and depends on whether or not the compression ratio (piston top dead center position) is variable. is not. Such piston stroke characteristics increase the piston stay time near the top dead center as compared with a conventional general engine in which the piston is connected to the crankshaft by a single connecting rod.

  Next, the structure around the combustion chamber 21 will be described with reference to FIG. The engine 1 is, for example, a gasoline engine. An intake port 22 and an exhaust port 23 are formed on each side surface of the cylinder head 20, and the intake port 22 and the exhaust port 23 open to the combustion chamber 21. The combustion chamber formed in this way is called a pent roof type. An intake valve 24 is provided at the open end of the intake port 23 to the combustion chamber 21, and an exhaust valve 25 is provided at the open end of the exhaust port 23 to the combustion chamber 21. A fuel injection valve 26 is provided below the open end of the intake port 22 so as to face the combustion chamber 21. A spark plug 27 is provided at the center of the ceiling of the combustion chamber 21, and an electrode 27 a of the spark plug 27 projects into the combustion chamber 21.

  In the engine 1, well-known stratified combustion is performed in the operating region on the low rotation speed and low load side. In the stratified combustion, a flammable air-fuel mixture is collected around the electrode 27a of the spark plug 27, and an air layer without fuel is formed around the mixture to burn. Specifically, the tumble control valve 28 provided near the upstream end of the intake port 22 is closed to generate a tumble flow (longitudinal vortex) in the combustion chamber 21 during the intake stroke. In order to preserve the generated tumble flow well, a shallow dish-like cavity 10 a is provided on the crown surface of the piston 10. The flow of air flowing in from the opening end of the intake port 22 along the wall of the cylinder 11 on the exhaust side in the combustion chamber 21 is reversed by the cavity 10a. This inverted air is directed along the wall of the cylinder 11 on the intake side toward the spark plug 27 on the ceiling of the combustion chamber 21 (see the arrow in the figure). Fuel is injected from the fuel injection valve 26 in the latter half of the compression stroke and mixed with air to form an air-fuel mixture. The injected fuel mixed with the air is guided to the electrode 27a of the spark plug 27 while being bundled together by the tumble flow. The ignition timing, which is the timing at which the spark plug 27 is ignited, is matched with the timing at which this block fuel reaches the electrode 27a of the spark plug 27, and a spark is applied to the electrode 27a of the spark plug 27 before the block fuel diffuses. Fly and ignite the bulk fuel. Since the air-fuel mixture is collected only around the electrode 27a of the spark plug 27 and there is air without fuel, that is, two layers are generated, so the combustion chamber 21 as a whole has an extremely lean air-fuel ratio. It becomes possible to burn in. On the other hand, on the high rotation speed side and the high load side, so-called homogeneous combustion is performed in which a homogeneous air-fuel mixture is formed in the entire combustion chamber 21 and burned. This concludes the description of the reciprocating engine having the variable compression ratio mechanism. In the present embodiment, the case where stratified combustion and homogeneous combustion are switched depending on the operating conditions will be described. However, the present embodiment is not limited to this case. Homogeneous combustion may be performed at the time of deceleration and reacceleration during high-speed constant speed driving, which will be described later, or stratified combustion may be performed.

  Returning to FIG. 4, the engine control system having the turbocharger 51 and the compression ratio variable mechanism will be outlined. The engine 1 shown in FIG. 4 does not have the compression ratio variable mechanism, but actually has the compression ratio variable mechanism shown in FIGS.

  In FIG. 2, the fuel injection valve 26 is provided directly facing the combustion chamber 21, and the stratified combustion is performed using the fuel injection valve 26 as described above. However, the present invention is not limited to this case, and the fuel injection valve 26 may be provided in the intake port 22 as shown in FIG. When the fuel injection valve 26 is provided in the intake port 22, stratified combustion cannot be performed, and therefore homogeneous combustion may be performed.

  The engine 1 is mounted on a vehicle. The engine 1 includes an intake passage 31 and an exhaust passage 41. The intake passage 31 includes an intake pipe 31a, an intake collector 31b, and an intake manifold 31c.

  The intake pipe 31a immediately upstream of the intake collector 31b is equipped with an electronically controlled throttle device that responds to the amount of depression of the accelerator pedal. The throttle device includes a throttle valve 32 and a motor (rotary electric machine) 33 that drives the throttle valve 32. The intake air is metered by the throttle valve 32 through the intake pipe 31a. The metered air is stored in the intake collector 31b, and is distributed and supplied from the intake collector 31b to the combustion chamber 21 of each cylinder via the intake manifold 31c. Although the embodiment is an electronically controlled throttle device, the throttle valve and the accelerator pedal may be connected by a wire.

  The exhaust passage 41 includes an exhaust manifold 41a into which exhaust gas from the combustion chamber 21 of each cylinder flows, and an exhaust pipe 41b connected to a collecting portion of the exhaust manifold 41a. The exhaust contains harmful three components of HC, CO and NOx. In order to purify all of these harmful components, manifold catalysts 42A and 42B are provided in the exhaust pipe 11b downstream of the turbine 52, and a main catalyst 43 is provided in the exhaust pipe 41b downstream thereof. The main catalyst 43 is provided, for example, under the floor of the vehicle. Each of these catalysts 42A, 42B, 43 is composed of, for example, a three-way catalyst. A muffler 49 is provided at the end of the exhaust pipe 41b.

  The engine 1 has a turbocharger 51. The reason why the variable compression ratio mechanism is adopted in the engine having the turbocharger 51 is to achieve both the fuel efficiency performance and the output / response performance of the engine. That is, in the case of an engine having a variable compression ratio mechanism, it is possible to improve thermal efficiency and improve fuel efficiency by setting a high target compression ratio on the low load side where knock is unlikely to occur. On the other hand, at the high load side with the supercharging region, the output / response performance can be improved by supercharging while suppressing the occurrence of knocking by setting the target compression ratio low.

  The turbocharger 51 includes a turbine 52 provided in the exhaust pipe 41b, a compressor 53 provided in the intake pipe 41a, and a shaft 54 connecting the turbine 52 and the compressor 53. The turbine 52 is rotated by the energy of the exhaust gas flowing through the exhaust pipe 41b, and drives a compressor 53 coaxial with the turbine 52. The compressor 53 compresses fresh air drawn through the air cleaner 57. Here, “fresh air” refers to air that does not contain EGR gas, which will be described later. Fresh air is also called air. The compressed air that is compressed and exceeds the atmospheric pressure is sent to the intake collector 31b. The target supercharging pressure can be obtained by operating the turbocharger 51 in the supercharging region on the high load side.

  The turbocharger 51 includes a bypass passage 54 that bypasses the turbine 52 and a normally closed waste gate valve 55 that opens and closes the bypass passage 54. The waste gate valve 55 is driven by a motor (rotating electric machine) 56. For example, when the actual boost pressure detected by the boost pressure sensor 95 becomes higher than the target boost pressure, the waste gate valve 55 is opened by driving the motor 56 and a part of the exhaust gas flowing into the turbine 52 is removed. The turbine 52 is bypassed to flow. As a result, the turbine rotation speed is lowered before opening the waste gate valve 55, and the compressor rotation speed coaxial with the turbine 52 is also lowered. When the compressor rotational speed decreases, the actual supercharging pressure decreases and matches the target supercharging pressure. The waste gate valve 55 is held at a timing at which the actual boost pressure matches the target boost pressure.

  The waste gate valve 55 and the throttle valve 32 are both devices for controlling the amount of fresh air. Here, in order to simplify the control method of the two valves 55 and 32, the fresh air amount is determined by the opening amount of the waste gate valve 55 in the supercharging region, and the fresh air amount is determined by the opening amount of the throttle valve 32 in the non-supercharging region. By controlling, the target air amount according to the operating condition is obtained.

  FIG. 5 is a characteristic diagram of the target boost pressure. A supercharging region is defined on the high load side of the entire operation region in which the horizontal axis is the engine rotation speed Ne and the vertical axis is the engine load. In FIG. 5, the broken line represents the boundary of whether to control with the opening degree of the throttle valve 32 or with the opening degree of the weight gate valve 55. The region on the high load side from the broken line is the supercharging region, and the region on the low load side from the broken line is the non-supercharging region. In the supercharging region, for example, the target supercharging pressure is set to be higher as the rotational speed is higher and the load is higher. The maximum supercharging pressure can be obtained at the high load side boundary in the supercharging region. In other words, the maximum supercharging pressure can be obtained with the accelerator pedal fully depressed. The maximum boost pressure that can be obtained with the accelerator pedal fully depressed is hereinafter referred to as “WOT set boost pressure”.

  The intake collector 31b includes a water-cooled intercooler 61. The reason for the water cooling type is to shorten the volume of the intake pipe 31a. In the case of an air-cooled intercooler, it is necessary to interpose the intake pipe 31a, and accordingly, the length of the intake pipe 31a is increased. On the other hand, since the water-cooled intercooler 61 can be provided inside the intake collector 31b, the length of the intake pipe 31a does not increase.

  The intercooler 61 is for cooling the air compressed by the compressor 53 with the cooling water flowing through the cooling water passage. Specifically, the intercooler 61 includes a cooling water passage and an air passage provided on the outer periphery thereof. The cooling water passage of the intercooler 61 and the air-cooled sub-radiator 62 are connected by cooling water passages 63 and 64. For example, the sub-radiator 62 is arranged in series with a radiator for cooling engine coolant so that traveling wind passes therethrough. The cooling water passage 63 is provided with a pump 65 for circulating the cooling water.

  In the sub-radiator 62, the warm cooling water carried from the intercooler 61 is cooled by traveling wind. The cooling water cooled by the sub radiator 62 is guided to the intercooler 61. In the intercooler 61, heat is exchanged between the air whose temperature has been increased by the air compression by the compressor 53 and the cooling water flowing through the cooling water passage, thereby cooling the air. The air whose temperature has been increased by the air compression by the compressor 53 is cooled by the intercooler 61, whereby the supercharging efficiency can be increased.

  Even in the gasoline engine 1 including the turbocharger 51, there is a demand for performing a large amount of EGR (exhaust gas recirculation) in order to suppress the occurrence of knocking in the supercharging region. In order to meet this requirement, the present embodiment includes a ropeless loop EGR device (hereinafter referred to as “low pressure EGR device”) 44. The low-pressure EGR device 44 includes an EGR passage 45, an EGR cooler 46 interposed in the EGR passage 45, an EGR valve 47 (for example, a butterfly valve) that opens and closes the EGR passage 45, and a motor (rotary electric machine) 48 that drives the EGR valve 47. Consists of.

  The EGR passage 45 is branched from an exhaust pipe downstream of the turbine 52, specifically, an exhaust pipe 41b between the manifold catalysts 42A and 42B and the main catalyst 43, and merges with an intake pipe 31a upstream of the compressor 53. Therefore, a part of the exhaust (a part of this exhaust is hereinafter referred to as “EGR gas”) flows through the EGR valve 47 due to the differential pressure between the exhaust pipe pressure downstream of the turbine and the intake pipe pressure upstream of the compressor. Since the differential pressure between the exhaust pipe pressure downstream of the turbine and the intake pipe pressure upstream of the compressor is as small as about 1 kPa, for example, it is called a low pressure EGR device.

  The EGR cooler 46 is provided in the EGR passage 45 upstream of the EGR valve 47. The EGR cooler 46 cools the EGR gas flowing through the EGR passage 45. For this reason, the cooled EGR gas flows through the EGR valve 47 in the EGR region.

  The EGR region is provided so as to partially overlap the above-described supercharging region. For this reason, the operation range is <1> supercharging region and EGR region, <2> supercharging region and non-EGR region, <3> non-supercharging region and EGR region, <4> non-supercharging region and non-EGR region. It is divided into four areas. Particularly in the region <1> described above, in the low-pressure EGR device 44, the EGR gas 45 flows into the EGR passage 45 with a small differential pressure (about 1 kPa) between the relatively low exhaust pipe pressure downstream of the turbine and the intake pipe pressure upstream of the compressor. So that it is not affected by supercharging pressure. That is, by adding a low-pressure EGR device 44 to the gasoline engine 1 provided with the turbocharger 51, a large amount of EGR gas is introduced (discharged) into the intake pipe 31a even during supercharging by the turbocharger 51. it can.

  The differential pressure may not be ensured in a region on the low rotation speed and low load side in the EGR region. In order to cope with this case, a differential pressure device is provided in the intake pipe 31a upstream of the joining portion of the EGR passage 45. The differential pressure device includes a butterfly valve 81 that opens and closes the intake pipe 31a upstream of the joining portion of the EGR passage 45, and a motor (rotating electrical machine) 82 that drives the butterfly valve 81.

  Furthermore, a bypass passage 71 that bypasses the compressor 53 is provided. The bypass passage 71 is provided with a recirculation valve 72 driven by a motor (rotating electric machine) 73. This valve 72 recirculates (recirculates) pressurized air confined in the intake pipe 31a from the throttle valve 32 to the compressor 53 upstream of the compressor 53 when the throttle valve 32 is closed for vehicle deceleration. It is for making it happen. On the other hand, when turbocharging is performed by the turbocharger 51 in an operating region other than when the vehicle is decelerating, the valve 72 is basically fully closed and the air upstream of the compressor 53 (including EGR gas) ) Are all guided to the compressor 53.

  An engine controller 91 is provided to control the compression ratio of the variable compression ratio mechanism by the compression ratio control actuator 17 described above. The engine controller 91 is configured as a computer unit including peripheral devices such as a microprocessor, ROM, and RAM. Signals from the air flow meter 92, the accelerator sensor 93, the crank angle sensor 94, and the supercharging pressure sensor 95 are input to the engine controller 91. Here, the air flow meter 92 detects the amount of air (mass flow rate) flowing into the intake pipe 31a. The accelerator sensor 93 detects an accelerator pedal depression amount (accelerator opening) APO and a change amount thereof. The crank angle sensor 94 detects the engine speed Ne. The supercharging pressure sensor 95 detects the pressure (actual supercharging pressure) of the intake collector 31b.

  The engine controller 91 calculates a target compression ratio according to the load and the rotational speed Ne at that time by referring to a map of the target compression ratio from the engine load and the rotational speed Ne after the engine warm-up is completed. Then, the control amount (drive amount to the compression ratio variable mechanism) given to the compression ratio control actuator 17 is controlled so that the calculated target compression ratio is obtained.

  The engine controller 91 controls the throttle valve 32, the EGR valve 47, the waste gate valve 55, and the recirculation valve 72 in addition to the fuel injection valve 26 and the spark plug 27. For example, the opening degree of the waste gate valve 55 is controlled so that the target boost pressure shown in FIG. 5 is obtained.

  Now, in the engine having the turbocharger 51 and the compression ratio variable mechanism, it is a problem to further improve the power performance and response performance of the engine. For this purpose, the air response (specifically, the air response speed) in the supercharging region is improved. Here, the “air response speed” refers to the degree of ease with which the amount of fresh air is supplied to the combustion chamber 21. To improve the “air response speed” means to quickly increase the amount of fresh air supplied to the combustion chamber 21. In this case, since the amount of fresh air is proportional to the supercharging pressure, “response speed of air” is synonymous with “response speed of supercharging pressure”. In the following, “supercharging pressure response speed” is used instead of “air response speed”.

  In order to meet the above problem (that is, to improve the response speed of the supercharging pressure), the turbocharger is reduced to a smaller capacity, for example, with a lighter motion, while maintaining the response speed of the compression ratio variable mechanism. Suppose that it has changed. Of course, since there are various methods for improving the response speed of the supercharging pressure, it is not limited to the case illustrated here. In the turbocharger 51, generally, when the engine rotational speed Ne increases, the exhaust flow rate increases, and the rotational speed of the turbine 52 increases early due to the increase in the exhaust flow rate. By using the turbocharger after the change exemplified above, the supercharging pressure can be increased earlier in the middle and high engine speed range than when using the turbocharger before the change, that is, the response speed of the supercharging pressure. Will be improved. In other words, the turbocharger after the change exemplified above does not work in the low rotational speed region but works in the middle and high rotational speed region. Here, “the turbocharger works” means that the actual supercharging pressure reaches the target supercharging pressure when the opening degree of the waste gate valve 55 is controlled. “The turbocharger does not work” means that the actual supercharging pressure does not reach the target supercharging pressure even if the opening degree of the waste gate valve 55 is controlled. In addition, when the entire rotation speed range is divided into three main parts, a low rotation speed range, a medium rotation speed range, and a high rotation speed range, the medium rotation speed range and the high rotation speed range are collectively referred to as “medium / high rotation speed range” Says.

  In the present embodiment, the entire rotational speed region is divided into two regions, a low rotational speed region and a medium / high rotational speed region. The turbocharger does not work in the low rotational speed region, and the turbocharger operates in the medium / high rotational speed region. Although described as working, it is not limited to this case. Which rotation speed range or which rotation speed range is appropriate to operate the turbocharger is a matter directly related to engine development. Therefore, the rotational speed range in which the turbocharger operates and the rotational speed range may differ depending on the engine specifications. However, due to the nature of the turbocharger, the turbocharger does not work in the low rotation speed region where the exhaust flow rate is not sufficient, and the turbocharger works in the high rotation speed region where sufficient exhaust flow rate can be obtained. It will be. Therefore, the engine rotation speed range is divided into two regions, the low rotation speed side and the high rotation side, with the predetermined rotation speed as a boundary, and the turbocharger does not work in the low rotation speed side region, and the high rotation speed side The present invention is applied when the turbocharger works in a region.

  As a result of improving the response speed of the supercharging pressure using the turbocharger after the change exemplified above, the difference from the response speed of the compression ratio of the compression ratio variable mechanism becomes large, and the following (1), ( The inventor has found that the problem 2) occurs. Hereinafter, the compression ratio of the compression ratio variable mechanism is also referred to as “actual compression ratio”.

  (1) The state where the vehicle is traveling at a constant speed on the highway is included in the middle and high rotation speed region where the turbocharger works. Following the previous vehicle (the vehicle traveling in front of the host vehicle), the front vehicle decelerated from the state of traveling at a constant speed on the highway, so the accelerator pedal is slightly returned to decelerate the host vehicle accordingly. Immediately after that, the previous vehicle accelerated, and in response to this, the accelerator pedal was depressed immediately to re-accelerate the vehicle and return to the original constant speed running. In this way, a driving scene in which the vehicle follows the front vehicle on a highway is often seen. Such a driving scene is hereinafter referred to as “when decelerating and reaccelerating while driving on a highway at a constant speed”.

  At the time of deceleration reacceleration while traveling on a highway at a constant speed, the target compression ratio shifts from a low side to a high side by deceleration, and returns from a high side to a low side by reacceleration immediately thereafter. In this way, the actual compression ratio follows the target compression ratio that changes from low to high and from high to low. On the other hand, the actual supercharging pressure decreases from the target supercharging pressure due to deceleration, but reaches the target supercharging pressure by reacceleration. Considering the movements of both the actual compression ratio and the actual supercharging pressure, the timing at which the actual compression ratio reaches the target compression ratio after reacceleration by reacceleration and the actual supercharging pressure is regenerated by reacceleration. Ideally, the timing to reach the target boost pressure after acceleration almost coincides.

  However, as a result of improving the response speed of the supercharging pressure using the turbocharger after the change exemplified above, the response speed of the supercharging pressure is too fast with respect to the response speed of the actual compression ratio. . More specifically, the actual supercharging pressure reaches the target boost pressure after reacceleration before the timing at which the actual compression ratio reaches the target compression ratio after reacceleration by reacceleration. Since the target compression ratio after re-acceleration is on the low side, the actual compression ratio reaches the target compression ratio after re-acceleration when the actual boost pressure reaches the target boost pressure after re-acceleration. Not that means that the actual compression ratio remains on the high side. Then, while the actual boost pressure remains high, a state where the actual compression ratio becomes higher than the target compression ratio after re-acceleration, that is, a high boost and high compression ratio state occurs, and knocking occurs. is there. When the turbocharger before the change was used, the timing at which the actual compression ratio reaches the target compression ratio after reacceleration due to reacceleration, and the target The timing to reach the supply pressure was almost the same. Nevertheless, the response speed of the compression ratio of the variable compression ratio mechanism remains unchanged, and the response speed of the supercharging pressure is improved by using the turbocharger after the change exemplified above. A new problem arises that knocking occurs during acceleration.

  (2) If the response speed of the compression ratio of the variable compression ratio mechanism (actual speed of response of the actual compression ratio) remains unchanged and the setting of the target compression ratio remains unchanged, the above problem (1) will be avoided. When the knock occurs, the ignition timing is retarded from the MBT. In the present embodiment, the engine controller 91 determines whether or not a knock has occurred based on a signal detected by the knock sensor 97 (see FIG. 4). For this reason, when knocking occurs, the ignition timing can be retarded by a predetermined amount from MBT. Alternatively, it is conceivable to limit the actual boost pressure to the target boost pressure after reacceleration until the actual compression ratio reaches the target compression ratio after reacceleration by reacceleration. However, in any of these measures, the response of the engine is greatly impaired at the time of deceleration and reacceleration during constant speed driving on the highway.

  In this case, the response speed of the compression ratio of the variable compression ratio mechanism mainly depends on the compression ratio control actuator 17. More specifically, when the compression ratio control actuator 17 is a motor, the response speed of the compression ratio of the variable compression ratio mechanism is determined by the response speed of the motor. For this reason, as a solution to the problem of (1), the compression ratio control actuator having a high response speed corresponding to the improvement of the response speed of the supercharging pressure using the turbocharger after the change exemplified above. It is possible to change to 17. However, changing to the compression ratio control actuator 17 having a high response speed increases the cost. Even if the turbocharger is changed, the cost can be increased. However, if the compression ratio control actuator 17 is changed, the cost is increased by double. Therefore, in this embodiment, it does not consider changing to the compression ratio control actuator 17 with a high response speed. That is, the response speed of the compression ratio of the variable compression ratio mechanism (the actual response speed of the compression ratio) remains unchanged.

  Therefore, in the first embodiment of the present invention, the problem of the above (1) is addressed by changing the setting of the target compression ratio in the middle and high rotation speed region where the turbocharger 51 works. This will be described with reference to FIGS. 6 and 7. FIG. 6 is a characteristic diagram of the target compression ratio of a comparative example (hereinafter simply referred to as “comparative example” in the first embodiment) with respect to the first embodiment. These are the characteristic figures of the target compression ratio of 1st Embodiment.

  The case of the comparative example will be described. As shown in FIG. 6, in the comparative example, the target compression ratio is set depending only on the engine load. That is, isocompression ratio lines for six typical target compression ratios are drawn substantially parallel to the left-right direction. Since knocking is unlikely to occur in the low load range, the target compression ratio is set high in order to improve the thermal efficiency of the engine. On the other hand, if the target compression ratio is increased in the high load region that is also the supercharging region, knocking is likely to occur. Therefore, the target compression ratio is set low in the high load region, and MBT having good thermal efficiency is used as the ignition timing. When the above six representative target compression ratios are predetermined values ε1, ε2, ε3, ε4, ε5, ε6, ε1 on the lowest load side is the highest and ε6 on the highest load side is the lowest. . The target compression ratio gradually decreases from ε1 to ε6 (ε1> ε2> ε3> ε4> ε5> ε6).

  Next, the case of the first embodiment will be described. As shown in FIG. 7, in the first embodiment, each of the six equal compression ratio lines is a straight line that goes diagonally downward to the right in the middle and high rotation speed region where the turbocharger 51 works. In this case, the target compression ratio is set to the lower side as the engine speed Ne becomes higher in the medium and high speed ranges.

  Here, the reason why the target compression ratio is set to the lower side as the engine rotational speed Ne becomes higher in the middle and high rotational speed range will be described. If you write in Fig. 6 how the driving point moves during deceleration and re-acceleration while driving at a constant speed on an expressway, the state of following the preceding vehicle and driving on the expressway at a constant speed is, for example, point A The target compression ratio at point A is assumed to be ε5. In this state, the front vehicle decelerates, and it is assumed that the vehicle is decelerated by returning the accelerator pedal accordingly. As a result, the engine load is reduced and the engine rotational speed is also reduced. Therefore, it is assumed that the operating point moves from point A to point B, and the target compression ratio at point B is ε1. Then, it is assumed that the vehicle is accelerated by depressing the accelerator pedal in an attempt to follow the vehicle that has accelerated immediately after deceleration. As a result, the engine load increases and the engine speed also increases, so that the operating point moves from point B to point C close to point A, and the target compression ratio at point C is ε5. In this way, the driving point moves from the upper right point A to the lower left point B and again moves from the lower left point B to the upper right point C at the time of deceleration and reacceleration while driving at a constant speed on the highway, that is, reciprocating diagonally. Will be.

  Considering such a movement of the operating point that reciprocates diagonally, in this embodiment, as shown in FIG. 7, the six equal compression ratio lines are in a direction intersecting the diagonal movement of the operating point. When the same operation as shown in FIG. 6 is performed, the target compression ratio at the points A and C becomes smaller than ε5 and the target compression ratio at the point B becomes smaller than ε1. Yes. More specifically, according to the characteristics shown in FIG. 7, the operating point is at point A before deceleration (before reacceleration), and the target compression ratio at point A is lower than ε5 (this target compression ratio). The compression ratio is “ε5 low”). Then, due to deceleration, the operating point moves from point A to point B, and the target compression ratio at point B becomes a target air-fuel ratio close to ε2 that is lower than ε1 (this target compression ratio is set to “ε1low”). Assume that the point moves from point B to point C, and the target compression ratio at point C is ε5 low. That is, in the comparative example, the target compression ratio has changed from ε5 → ε1 → ε5, but according to the characteristics of FIG. 7, in the present embodiment, the target compression ratio changes from ε5low → ε1low → ε5low. As a result, the amount of change (ε5 low to ε1 low) in the target compression ratio before and after the deceleration reacceleration of the present embodiment is smaller than the amount of change (ε5 to ε1) in the target compression ratio before and after the deceleration reacceleration in the comparative example. .

  The reason for this will be described next. The distance (the distance between A and B and the distance between B and C) in which the operating point moves before and after deceleration and reacceleration in FIG. 6 and FIG. 7 is the same. Further, the interval (shortest distance) between two equal compression ratio lines adjacent in FIG. 6 and FIG. 7 is the same. Then, even if the operating point moves through the same length, the target compression ratio change margin (ε5low to ε1low) when the operating point moves so as to be orthogonal to the isocompression ratio line is the same compression ratio line. This is because the change margin (ε5 to ε1) when the operating point moves so as not to be orthogonal to the operating point becomes smaller.

  In this way, if the change margin of the target compression ratio before and after deceleration and re-acceleration is small, even if the response speed of the actual compression ratio is the same, the time for the operating point to reciprocate is shortened. The timing to reach the target compression ratio after reacceleration is earlier than in the comparative example. As a result, the timing at which the actual compression ratio reaches the target compression ratio after reacceleration due to reacceleration and the timing at which the actual supercharging pressure reaches the target boost pressure after reacceleration due to reacceleration are substantially matched. be able to. The timing at which the actual compression ratio reaches the target compression ratio after reacceleration due to reacceleration and the timing at which the actual boost pressure reaches the target boost pressure after reacceleration due to reacceleration are almost the same. As a result, a high supercharging and high compression ratio state does not occur. The occurrence of knocking is avoided if a high supercharging and high compression ratio condition does not occur. By setting the target compression ratio on the side that becomes lower as the engine rotation speed becomes higher in the middle and high rotation speed range, the response speed of the actual compression ratio remains unchanged and the modified turbocharger exemplified above is used. It is possible to avoid the occurrence of knocking at the time of deceleration and re-acceleration while driving on a highway at a constant speed.

  FIG. 6 shows a case where the target compression ratio changes as ε5 → ε1 → ε5 during deceleration and reacceleration during traveling at a constant speed on the highway. However, the present invention is not limited to this case. For example, the target compression ratio may change as ε6 → ε1 → ε6, ε4 → ε1 → ε4, ε3 → ε1 → ε3, and ε2 → ε1 → ε2 during deceleration and reacceleration during constant speed driving on a highway. Similarly, it is possible that the target compression ratio does not increase to ε1 due to deceleration, and changes as ε5 → ε2 → ε5, ε5 → ε3 → ε5, and ε5 → ε4 → ε5. Further, the target compression ratio before re-acceleration may be different from the target compression ratio after re-acceleration, as in the case where the target compression ratio changes from ε5 → ε1 → ε6 and ε5 → ε1 → ε4. Furthermore, it may be limited to re-acceleration while driving at a constant speed on a general road, and only during deceleration and re-acceleration while driving at a constant speed on an expressway.

  The characteristic shown in FIG. 7 is an example, and the slope of the isocompression ratio line in the middle and high rotation speed region may vary depending on the engine specifications. In FIG. 7, the movement of the operating point happens to be almost orthogonal to the isocompression ratio line, but this is not a limitation.

  Next, with reference to FIG. 8A, a method for setting the target compression ratio in the middle and high rotation speed range will be described in detail. As described above, the problem (1) occurs when the response speed of the supercharging pressure is too high with respect to the response speed of the compression ratio (actual compression ratio) of the variable compression ratio mechanism. Rethink with response time. Here again, consider the case of deceleration and re-acceleration while driving on a highway at a constant speed. In this case, as the response time of the supercharging pressure, the time until the actual supercharging pressure reaches the target supercharging pressure after the reacceleration by reacceleration is taken. On the other hand, as the response time of the actual compression ratio, the time until the actual compression ratio reaches the target compression ratio after the reacceleration by reacceleration is taken. For this reason, in FIG. 8A, the horizontal axis represents the engine rotation speed Ne, and the vertical axis represents time (response time).

  The response time of the supercharging pressure becomes the longest when reaccelerating the deceleration while driving at a constant highway speed when the accelerator pedal is fully depressed during the reacceleration. As described above with reference to FIG. 5, the maximum supercharging pressure that can be obtained with the accelerator pedal fully depressed to re-accelerate is the WOT set supercharging pressure. Here, it is assumed that the WOT set supercharging pressure is a constant value. The time it takes for the actual boost pressure to reach the WOT set boost pressure during deceleration and re-acceleration while driving at a constant speed on the highway is given by a downwardly convex curve downward as shown by the dashed line in FIG. 8A. It is done. In this way, a downwardly convex curve with a lower right angle means that more exhaust flows into the turbine 52 as the engine rotational speed Ne increases in the middle and high rotational speed range, and the time until the WOT set supercharging pressure is reached is shorter. It is to become.

  Here, the case of the WOT set supercharging pressure will be described, but the present invention is not limited to the supercharging pressure. For example, when the maximum engine torque obtained with the accelerator pedal fully depressed to re-accelerate is defined as “WOT set torque”, this WOT set torque is determined in advance. Instead of the time until the actual boost pressure reaches the target boost pressure after reacceleration due to reacceleration, the time until the actual engine torque reaches this WOT set torque due to reacceleration is used. It's okay.

  Further, in the present embodiment, the case where the response time of the actual supercharging pressure becomes the longest, that is, the time until the actual supercharging pressure reaches the WOT set supercharging pressure by re-acceleration will be described. I can't. For example, the boost pressure obtained when the accelerator pedal is depressed to a predetermined opening that is not full to re-accelerate is set to “set boost pressure”. The time until the supercharging pressure reaches the set supercharging pressure may be adopted. Similarly, the engine torque obtained when the accelerator pedal is depressed to a predetermined opening that is not full to re-accelerate is set torque, and the actual engine torque during deceleration re-acceleration during high-speed constant-speed driving May be the time until the set torque is reached.

  On the other hand, the time until the actual compression ratio reaches the target compression ratio after re-acceleration is given as shown by the broken line in FIG. 8A. To explain this, assuming that the target compression ratio before deceleration is ε6 during deceleration and reacceleration during constant speed driving on the highway, the target compression ratio is increased to ε1 by performing deceleration to return the accelerator pedal to full, It is assumed that the target compression ratio has returned to ε6 by reacceleration. Before the re-acceleration and after the re-acceleration, the target compression ratio is set to the same ε6, the amount of returning the accelerator pedal during deceleration is reduced, and the target compression ratio is increased to ε2 by deceleration. Similarly, the amount by which the accelerator pedal is returned during deceleration under the same conditions is reduced, and the target compression ratio is increased to ε2, ε3, ε4, and ε5 by deceleration. Thus, while the target compression ratio is set to the same ε6 before and after the deceleration, by changing the return amount of the accelerator pedal at the time of the deceleration, five times indicated by broken lines in FIG. 8A can be obtained.

  That is, the uppermost broken line in FIG. 8A shows the time until the actual compression ratio reaches ε6 by reacceleration when the target compression ratio changes from ε6 → ε1 → ε6 (in FIG. 8A, “ε6 → ε1 → abbreviated as “time to change to ε6”). The second broken line from the top in FIG. 8A indicates the time until the actual compression ratio reaches ε6 by reacceleration when the target compression ratio changes from ε6 → ε2 → ε6 (in FIG. 8A, “ε6 → ε2 → abbreviated as “time to change as ε6”.). The third broken line from the top in FIG. 8A indicates the time until the actual compression ratio reaches ε6 by reacceleration when the target compression ratio changes from ε6 → ε3 → ε6 (in FIG. 8A, “ε6 → ε3 → abbreviated as “time to change as ε6”.). The fourth broken line from the top in FIG. 8A shows the time until the actual compression ratio reaches ε6 by reacceleration when the target compression ratio changes from ε6 → ε4 → ε6 (in FIG. 8A, “ε6 → ε4 → abbreviated as “time to change as ε6”.). The broken line at the bottom in FIG. 8A shows the time until the actual compression ratio reaches ε6 by reacceleration when the target compression ratio changes from ε6 → ε5 → ε6 (in FIG. 8A, “ε6 → ε5 → ε6”). And “changing time”.

  The method of handling the time until the actual compression ratio reaches the target compression ratio after re-acceleration is not limited to the case of FIG. 8A. For example, it may be the case of FIG. 8B. To explain this, assuming that the target compression ratio before reacceleration was ε6 during deceleration reacceleration while driving at a constant speed on an expressway, the target compression ratio is increased to ε1 by performing deceleration to return the accelerator pedal to full, It is assumed that the target compression ratio has returned to ε6 by reacceleration. Next, the target compression ratio before reacceleration and after reacceleration is set to the same ε5, and at the time of deceleration, the accelerator pedal is fully returned. Similarly, the target compression ratios before reacceleration and after reacceleration are set to the same ε4, ε3, and ε2, respectively, and the accelerator pedal is returned to full when decelerating. Thus, by setting the target compression ratio to the same ε6, 5, 4, 3, 2 before and after deceleration, and returning the accelerator pedal to the full during deceleration, the five response times indicated by the broken lines in FIG. 8B can get.

  That is, the uppermost broken line in FIG. 8B indicates the time until the actual compression ratio reaches ε6 by reacceleration when the target compression ratio changes from ε6 → ε1 → ε6 (in FIG. 8B, “ε6 → ε1 → abbreviated as “time to change to ε6”). The second broken line from the top in FIG. 8B shows the time until the actual compression ratio reaches ε5 by reacceleration when the target compression ratio changes from ε5 → ε1 → ε3 (in FIG. 8B, “ε5 → ε1 → abbreviated as “time to change as ε5”). The third broken line from the top in FIG. 8B shows the time until the actual compression ratio reaches ε4 by reacceleration when the target compression ratio changes from ε4 → ε1 → ε4 (in FIG. 8B, “ε4 → ε1 → abbreviated as “time to change as ε4”). The fourth broken line from the top in FIG. 8B shows the time until the actual compression ratio reaches ε3 by reacceleration when the target compression ratio changes from ε3 → ε1 → ε3 (in FIG. 8B, “ε3 → ε1 → abbreviated as “time to change as ε3”). The bottom broken line in FIG. 8B indicates the time until the actual compression ratio reaches ε2 by reacceleration when the target compression ratio changes from ε2 → ε1 → ε2 (in FIG. 8B, “ε2 → ε1 → ε2”). And “changing time”.

  As shown in FIGS. 8A and 8B, by using the turbocharger after the change exemplified above, the time until the actual supercharging pressure reaches the WOT set supercharging pressure in the medium and high rotation speed range can be obtained. The time until the actual compression ratio reaches the target compression ratio after re-acceleration is longer. As a result, it can be seen that, during deceleration and re-acceleration during traveling on a highway at a constant speed, a state of high supercharging and a high compression ratio occurs, and knocking is likely to occur. Therefore, in order to prevent the time until the actual compression ratio reaches the target compression ratio after reacceleration from exceeding the time until the actual supercharging pressure reaches the WOT set supercharging pressure, The time until the compression ratio reaches the target compression ratio after re-acceleration is set so as not to exceed the one-dot chain line. For this purpose, the target compression ratio in the medium-high rotation speed region is set so that the time until the actual compression ratio reaches the target compression ratio after reacceleration follows the solid line. Here, the solid line follows the one-dot chain line in the middle and high rotation speed range. To make the time until the actual compression ratio reaches the target compression ratio after reacceleration follow the solid line, the target compression ratio to which the actual compression ratio reaches due to reacceleration is lower than in the comparative example Let me do it. In addition, it is necessary to shorten the time until the actual compression ratio reaches the target compression ratio after the re-acceleration as the rotational speed in the medium-high rotational speed range increases. Correspondingly, the higher the rotational speed in the medium-high rotational speed region, the lower the target compression ratio to which the actual compression ratio reaches by reacceleration than in the comparative example. FIG. 7 shows the result of setting the target compression ratio in the medium / high rotation speed region in this way, and the target compression ratio is set to the side where the engine rotation speed Ne becomes higher in the medium / high rotation speed region.

  Further explanation will be given. In the comparative example, at the time of deceleration reacceleration while driving on a highway at a constant speed, according to the characteristics of FIG. 6, the target compression ratio changed from ε5 to ε1 by deceleration and then changed to ε5 by reacceleration. . On the other hand, under the same operating conditions as in the comparative example, in the present embodiment, according to the characteristics shown in FIG. 7, the operating point is at point A before deceleration (before reacceleration), and the target compression ratio at point A is ε5 low. Then, the operating point moves from point A to point B due to deceleration, the target compression ratio at point B becomes ε1 low, the operating point moves from point B to point C due to reacceleration, and the target compression ratio at point C becomes ε5 low. . In the comparative example, the target compression ratio has changed from ε5 → ε1 → ε5, but according to the characteristics of FIG. 7, it changes from ε5low → ε1low → ε5low under the same operating conditions as in the comparative example. As a result, the amount of change (ε5 low to ε1 low) in the target compression ratio before and after the deceleration reacceleration of the present embodiment is smaller than the amount of change (ε5 to ε1) in the target compression ratio before and after the deceleration reacceleration in the comparative example. As mentioned above. If the amount of change in the target compression ratio before and after the deceleration reacceleration in this embodiment is smaller, the time until the actual compression ratio reaches the target compression ratio after the reacceleration is shortened accordingly. Since the response speed of the actual compression ratio remains the same as the comparative example, the actual compression ratio reaches the target compression ratio after re-acceleration by setting the target compression ratio in the middle and high rotation speed range lower than the comparative example. The time to do it is shortened. As a result, the time required for the actual compression ratio to reach the target compression ratio after reacceleration during deceleration and reacceleration during constant speed driving on the highway is the time until the actual supercharging pressure reaches the WOT set supercharging pressure. You can avoid exceeding the time.

  According to the characteristic of the target compression ratio shown in FIG. 7 set in this way, even at the time of deceleration and re-acceleration while driving on a highway at a constant speed, a state of high supercharging and a high compression ratio does not occur. Therefore, the occurrence of knock can be avoided. From this, at the time of deceleration reacceleration while driving on a highway at a constant speed, the ignition timing is retarded from MBT, or the supercharging pressure is limited until the actual compression ratio reaches the target compression ratio after reacceleration. Therefore, the engine response can be improved.

  This control performed by the engine controller 91 will be described with reference to the flowchart of FIG. The flow in FIG. 9 is for setting the target compression ratio, and is executed at regular intervals (for example, every 10 ms).

  In step 1, the engine load and the engine speed Ne are read. The engine rotation speed Ne is detected by a crank angle sensor 94. As the engine load, for example, an accelerator pedal depression amount detected by the accelerator sensor 93 may be used.

  In step 2, a target compression ratio is calculated by searching a map having the contents shown in FIG. 7 from the read engine load and engine speed Ne. As shown in FIG. 7, in the present embodiment, the target compression ratio in the middle and high rotation speed region is set lower than in the comparative example.

  The target compression ratio calculated in this way is stored in a memory.

  In a flow not shown, the control amount (drive amount to the compression ratio variable mechanism) given to the compression ratio control actuator 16 is controlled so that the target compression ratio calculated in this way is obtained.

  Here, the effect of this embodiment is demonstrated.

  In the present embodiment, the turbocharger 51 (supercharger), the compression ratio variable mechanism, the crank angle sensor 94 (rotational speed detection means), the accelerator sensor 93 (load detection means), and the target compression ratio setting means. (91) and control means (91). The compression ratio variable mechanism can change the compression ratio of the engine. The crank angle sensor 94 detects the rotational speed of the engine. The accelerator sensor 93 detects the engine load. Based on the detected engine speed and the detected engine load, the target compression ratio setting means (91) sets the target compression ratio to a low speed in the middle and high speed range (high speed side area) even with the same load. Set to the lower side of the area (low speed side area). The control means (91) controls the compression ratio variable mechanism so that the set target compression ratio is obtained. In the present embodiment, the target compression ratio is set lower than that in the comparative example before and after the deceleration reacceleration during deceleration reacceleration during constant speed driving on the highway. Since the target compression ratio is set lower before and after the deceleration reacceleration, the target compression ratio change amount (ε5low to ε1low) before and after the deceleration reacceleration is the target compression ratio before and after the deceleration reacceleration of the comparative example. It becomes smaller than the change allowance (ε5 to ε1). If the amount of change in the target compression ratio before and after deceleration and re-acceleration is small, the time during which the operating point moves through the amount of change in the target compression ratio is shortened even if the response speed of the actual compression ratio is the same. From this, the timing when the actual compression ratio reaches the target compression ratio after re-acceleration during deceleration and re-acceleration while driving at constant speed on the highway, and the timing when the actual boost pressure reaches the target boost pressure after re-acceleration Can be equivalent. Thus, the actual compression ratio does not reach the target compression ratio after reacceleration at the timing when the actual supercharging pressure reaches the target boost pressure after reacceleration. Since a state of high supercharging and a high compression ratio does not occur, occurrence of knocking can be suppressed. If the occurrence of knocking can be suppressed, there is no need to retard the ignition timing or limit the actual supercharging pressure, so the engine torque or response in the supercharging region during deceleration reacceleration during constant speed driving on a highway. Can be improved.

  In the present embodiment, when the target compression ratio setting means (91) is decelerating and reaccelerating during constant speed driving on the highway (when decelerating and reaccelerating from a state where the vehicle is traveling at a constant speed in the region on the high rotational speed side), The target compression ratio in the middle and high rotation speed range (target compression ratio in the high rotation speed side area) is set as follows. That is, the time until the actual boost pressure reaches the WOT set boost pressure (set boost pressure) exceeds the time until the compression ratio of the compression ratio variable mechanism reaches the target compression ratio after re-acceleration. The target compression ratio in the middle and high rotation speed range is set to prevent this. As a result, the target compression ratio in the middle and high rotation speed range can be precisely set in accordance with the response performance requirement imposed on the turbocharger 51, so that the engine torque or response in the supercharge range can be reduced. It can be further improved.

  In this embodiment, when the total rotational speed region is divided into three regions of a low rotational speed region, a medium rotational speed region, and a high rotational speed region, the region on the high rotational speed side is the medium high rotational speed region and the low rotational speed region. The region on the speed side is the low rotational speed region. As a result, the working area of the turbocharger is expanded to the middle and high rotational speed range. Even when the deceleration and reacceleration during high speed constant speed traveling is included in the medium and high rotation speed range, it is possible to suppress the occurrence of knocking during deceleration and reacceleration during highway constant speed traveling.

(Second Embodiment)
FIG. 10 is a characteristic diagram of the target compression ratio of the second embodiment, which replaces FIG. 7 of the first embodiment. The same parts as those in FIG. 7 of the first embodiment are denoted by the same reference numerals.

  As shown in FIG. 10, in the second embodiment, the target compression ratio in the high rotation speed region is set higher than in the first embodiment. This takes into account the engine characteristics in the high rotational speed region. That is, as an engine characteristic, the knock margin becomes small in the high rotation speed range. Here, the knock margin is the interval (unit is deg) between the ignition timing of the knock limit (or the trace knock ignition timing) and MBT (Minimum advance for Best Torque). Moreover, the sensitivity of knock detection by the knock sensor 97 is not so high in the high rotation speed range. In other words, in a high rotational speed region, it becomes possible to operate at a place where the combustion is good near the MBT. In other words, the high rotational speed range has engine characteristics that knocking is less likely to occur without lowering the target compression ratio because the response deterioration margin due to knock is smaller than in the middle rotational speed range. . However, even if the target compression ratio can be set high in the high rotation speed region, it cannot be set as high as shown in FIG.

  The characteristics of the target compression ratio in the high rotational speed range vary depending on the engine specifications and are determined by conformance.

  In the second embodiment, the target compression ratio setting means (91) sets the target compression ratio to a higher side as the rotational speed of the engine becomes higher in a higher rotational speed range, while ensuring sufficient engine response. It is possible to improve fuel efficiency in the high rotation speed range.

(Third embodiment)
FIG. 11 is a characteristic diagram of a target compression ratio of a comparative example (hereinafter simply referred to as “comparative example” in the third embodiment) with respect to the third embodiment. FIG. 12 is a characteristic diagram of the target compression ratio of the third embodiment, which replaces FIG. 7 of the first embodiment. The same parts as those in FIG. 7 of the first embodiment are denoted by the same reference numerals.

  When the compression ratio at which the maximum combustion pressure Pmax is obtained is defined as “Pmax limit compression ratio”, the target compression ratio is exceeded when the operating point at which this Pmax limit compression ratio is reached is limited to the high rotational speed and high load side. Cannot be set. This is due to the following reason. That is, when a target compression ratio higher than the Pmax limit compression ratio is set on the high rotation speed and high load side from the operating point at which the Pmax limit compression ratio is set, the operation speed at the high rotation speed and high load side from the operation point at which the Pmax limit compression ratio is set. When the operating point comes, a combustion pressure exceeding the maximum combustion pressure Pmax is generated. If combustion pressure exceeding the maximum combustion pressure Pmax is generated, it is not preferable in terms of engine durability.

  The operating point at which the Pmax limit compression ratio is obtained differs depending on the engine specifications. For this reason, there may be an engine that cannot be applied with the characteristics of the target compression ratio shown in FIG. For example, in the comparative example, the operating point at which the Pmax limit compression ratio is reached is in the middle / high rotation speed region and the middle / high load region as shown by the broken line in FIG. Here, when the entire load area is divided into three main parts, a low load area, a medium load area, and a high load area, the medium load area and the high load area are collectively referred to as a “medium / high load area”. In the first embodiment, the operating point at which the Pmax limit compression ratio is reached in the medium and high rotation speed range and the medium and high load range, but in the third embodiment (including the comparative example), the medium and high rotation speed range and the medium and high load range. Therefore, there is an operating point at which the Pmax limit compression ratio is reached.

  Thus, when the engine specifications of the third embodiment (including the comparative example) and the first embodiment are different, the characteristics of the target compression ratio shown in FIG. 7 are applied to the engine of the comparative example as they are. As shown in FIG. That is, in the engine of the comparative example, as shown in FIG. 11, the equal compression ratio line of ε6 runs beyond the operating point where the Pmax limit compression ratio is reached to the high rotation speed and high load side. In other words, a target compression ratio higher than the Pmax limit compression ratio is set in a region where the operating point at which the Pmax limit compression ratio is reached exceeds the high rotation speed and high load side. As a result, when the operating point comes to the high rotational speed and high load side from the operating point at which the Pmax limit compression ratio is reached, a situation that is undesirable in terms of engine durability as described above occurs. In order to avoid such a situation, a region on the high rotational speed and high load side from the operating point at which the Pmax limit compression ratio is set is set as a target compression ratio setting prohibited region, and the target compression ratio is not set in this setting prohibited region.

  Therefore, in the third embodiment, the target compression ratio in the medium and high rotation speed region is set so that the operating point that becomes the Pmax limit compression ratio is the limit and the limit is not exceeded on the high rotation speed and high load side. Specifically, as shown in FIG. 12, the constant compression ratio line of ε6 is set so as to be closer to the low rotational speed and low load side than the operating point at which the Pmax limit compression ratio is reached. In accordance with the change in the ε6 isocompression ratio line, the ε5 isocompression ratio line is set to be moved to a lower rotational speed and lower load side than the ε6 isocompression ratio line. As a result, the set range of all target compression ratios is within the region at the low rotational speed and low load side from the operating point at which the Pmax limit compression ratio is achieved.

  In FIG. 11 and FIG. 12, the operating point that becomes the Pmax limit compression ratio is indicated by a straight line (see the broken line) in the middle and high rotation speed region and the middle and high load region, but this is not a limitation. The straight lines shown in FIGS. 11 and 12 are only shown as models. Actually, the operating point at the Pmax limit compression ratio may move in the load direction (vertical direction in FIGS. 11 and 12) depending on the engine specifications. Further, the operating point at which the Pmax limit compression is performed can be an upwardly convex curve or a downwardly convex curve depending on engine specifications.

  In the third embodiment, the target compression ratio setting means (91) further limits the operating point at which the Pmax limit compression ratio is reached as a limit, so that this limit is not exceeded on the high rotation speed and high load side. Set the target compression ratio in the rotation speed range and high rotation speed range). As a result, the engine function can be ensured because the target compression ratio higher than the Pmax limit compression ratio is not set on the higher rotation speed and higher load side than the operating point at which the Pmax limit compression ratio is achieved.

(Fourth embodiment)
Also in the fourth embodiment, it is assumed that the middle and high rotation speed range includes the time of deceleration and reacceleration during traveling at a constant highway speed. Now, during deceleration and re-acceleration while driving on a highway at a constant speed, the actual environmental condition parameters in the medium and high rotation speed range are different from the environmental condition parameters when the target compression ratio is set (adapted) in the medium and high rotation speed range. It is possible that Here, as environmental condition parameters,
<1> Intake air temperature (intake air temperature),
<2> Atmospheric temperature,
<3> Outlet temperature of the intercooler 61 (simply referred to as “intercooler temperature”),
<4> Engine coolant temperature,
<5> Engine lubricating oil temperature (oil temperature)
<6> Atmospheric pressure,
<7> Air humidity,
<8> Combustion chamber temperature,
Can be mentioned. <1> to <8> are environmental condition parameters that affect knocking. The combustion chamber temperature of <8> is a superordinate concept of the wall surface temperature in the combustion chamber 21 or the piston 10 temperature. In other words, the combustion chamber temperature can be substituted by the wall surface temperature in the combustion chamber 21 or the piston 10 temperature.

  When the environmental condition parameters <1> to <8> are arranged, the temperatures <1> to <5> and the pressure <6> all affect the combustion chamber temperature <8>. The combustion chamber temperature affects the knock. Therefore, here, the combustion chamber temperature of <8> is first taken up as a representative.

  As described above, the combustion chamber temperature when the actual combustion chamber temperature in the medium and high rotation speed range sets (adapts) the target compression ratio in the medium and high rotation speed range during deceleration and reacceleration while driving at a constant speed on the highway. In contrast, knocking may occur when the temperature of the combustion chamber is higher than that at the time of adaptation. On the other hand, during deceleration and reacceleration while driving on a highway at constant speed, the actual combustion chamber temperature in the medium and high rotation speed range is different from the combustion chamber temperature when the target compression ratio is set (adapted) in the medium and high rotation speed range, When the temperature is lower than the combustion chamber temperature at the time of adaptation, the fuel efficiency is deteriorated.

  Therefore, in the fourth embodiment, the target compression ratio in the medium / high rotation speed region is corrected based on the actual combustion chamber temperature in the medium / high rotation speed region. Specifically, depending on the displacement (difference or ratio) between the actual combustion chamber temperature in the medium-high rotation speed range and the target compression ratio in the medium-high rotation speed range (adaptation), the medium-high rotation speed Correct the target compression ratio in the speed range to lower or increase the target compression ratio.

  This will be specifically described. The timing at which the actual compression ratio reaches the target compression ratio after reacceleration and the timing at which the actual supercharging pressure reaches the target supercharging pressure after reacceleration are the same during deceleration and reacceleration while driving on a highway at constant speed It was mentioned above that it can be. If the combustion chamber temperature at this timing is the combustion chamber temperature at the time of adaptation, knocking does not occur. On the other hand, it is assumed that the actual combustion chamber temperature in the medium-high rotation speed region is higher by a predetermined value ΔT1 than the combustion chamber temperature when the target compression ratio in the medium-high rotation speed region is set (adapted). In this case, the combustion chamber temperature is too high at the above timing, and knocking may occur. In this case, in order to prevent the combustion chamber temperature from being excessively higher than that at the time of adaptation, it is necessary to reduce the target compression ratio in the middle and high rotation speed range. This is because if the target compression ratio is lowered, the thermal efficiency of the engine is deteriorated, and therefore the combustion chamber temperature is lowered before the target compression ratio is lowered.

  Therefore, in the fourth embodiment, a difference ΔT1 between the actual combustion chamber temperature in the medium and high rotation speed range and the combustion chamber temperature when the target compression ratio in the medium and high rotation speed range is adapted is calculated, and according to the difference ΔT1. Correction is made to reduce the target compression ratio in the middle and high rotation speed range. Here, the actual combustion chamber temperature is detected by a temperature sensor. For example, the wall surface temperature or piston 10 temperature in the combustion chamber 21 is detected by a temperature sensor, and the detected temperature is used as the actual combustion chamber temperature. As a result, even when the actual combustion chamber temperature is higher than the combustion chamber temperature when the target compression ratio is set (adapted) in the medium to high rotation speed region during deceleration and reacceleration during constant speed driving on the highway, Occurrence can be suppressed.

  On the other hand, it is assumed that the actual combustion chamber temperature is lower by a predetermined value ΔT2 than the combustion chamber temperature when the target compression ratio is set (adapted) in the middle and high rotation speed range during deceleration and reacceleration during constant speed driving on the highway. In this case, the combustion chamber temperature is too low at the above timing, and the thermal efficiency of the engine decreases. If the thermal efficiency of the engine decreases, the fuel efficiency will decrease accordingly. In this case, in order to prevent the combustion chamber temperature from being too low as compared with that at the time of adaptation, it is necessary to increase the target compression ratio in the middle and high rotation speed range. This is because if the target compression ratio is increased, the thermal efficiency of the engine is improved, so that the temperature of the combustion chamber rises before the target compression ratio is increased.

  Therefore, in the fourth embodiment, a difference ΔT2 between the actual combustion chamber temperature in the medium and high rotation speed range and the combustion chamber temperature when the target compression ratio in the medium and high rotation speed range is adapted is calculated, and according to the difference ΔT2. Correct the target compression ratio in the middle and high rotation speed range to the higher side. This makes it possible to reduce fuel consumption even when the actual combustion chamber temperature is lower than the combustion chamber temperature when the target compression ratio is set (adapted) in the medium to high rotational speed range during deceleration and reacceleration during constant speed driving on the highway. Can be better.

  Next, the atmospheric humidity of the above <7> (hereinafter referred to as “atmospheric humidity”) affects knocking. The relationship between atmospheric humidity and knock is as follows. That is, the high atmospheric humidity means that there is a lot of moisture in the atmosphere, and since the combustion is suppressed by the moisture in the atmosphere, the occurrence of knocking is suppressed. On the other hand, when the atmospheric humidity is low, the moisture in the atmosphere is small, and combustion is promoted accordingly and knocking is likely to occur.

  As described above, when reaccelerating the deceleration while driving on a highway at a constant speed, the actual atmospheric humidity in the medium and high rotation speed range is different from the atmospheric humidity when the target compression ratio in the medium and high rotation speed range is set (adapted). If the humidity is lower than the atmospheric humidity at the time of adaptation, knocking may occur. On the other hand, during deceleration and reacceleration while driving on a highway at a constant speed, the actual atmospheric humidity in the medium / high rotation speed range is different from the atmospheric humidity when the target compression ratio is set (adapted) in the medium / high rotation speed range. If it is higher than the atmospheric humidity, the fuel efficiency will deteriorate.

  Therefore, in the fourth embodiment, the target compression ratio in the medium / high rotation speed region is corrected based on the actual atmospheric humidity in the medium / high rotation speed region. Specifically, depending on the atmospheric humidity deviation (difference or ratio) when setting (adapting) the actual atmospheric humidity in the medium-high rotation speed range and the target compression ratio in the medium-high rotation speed range, The target compression ratio is corrected to be lower or higher.

  This will be specifically described. The timing at which the actual compression ratio reaches the target compression ratio after reacceleration and the timing at which the actual supercharging pressure reaches the target supercharging pressure after reacceleration are the same during deceleration and reacceleration while driving on a highway at constant speed It was mentioned above that it can be. If the atmospheric humidity at this timing is the atmospheric humidity at the time of adaptation, knocking does not occur. On the other hand, it is assumed that the actual atmospheric humidity in the medium / high rotation speed region is lower by a predetermined value ΔH1 than the atmospheric humidity when the target compression ratio in the medium / high rotation speed region is set (adapted). In this case, the atmospheric humidity is too low at the above timing, and knocking may occur. In this case, in order to prevent knocking from being easily generated because the atmospheric humidity is too lower than that at the time of adaptation, the target compression ratio in the middle and high rotation speed range should be reduced. This is because if the target compression ratio is lowered, the thermal efficiency of the engine is deteriorated, and the combustion chamber temperature is lowered before the target compression ratio is lowered.

  Therefore, in the fourth embodiment, the absolute value of the difference ΔH1 between the actual atmospheric humidity in the medium / high rotation speed region and the atmospheric humidity when the target compression ratio in the medium / high rotation speed region is adapted is calculated, and the absolute value of the difference ΔH1 is calculated. Depending on the value, the target compression ratio in the middle and high rotation speed range is corrected to the side that decreases. Here, the actual atmospheric humidity is detected by a humidity sensor. This prevents knocking even when the actual atmospheric humidity is lower than the atmospheric humidity when the target compression ratio is set (adapted) in the medium and high rotation speed range during deceleration and reacceleration during constant speed driving on the highway. Can be suppressed.

  On the other hand, it is assumed that the actual atmospheric humidity is higher by a predetermined value ΔH2 than the atmospheric humidity when the target compression ratio is set (adapted) in the middle and high rotation speed region during deceleration and reacceleration during constant speed driving on the highway. At this time, it is resistant to knocking as much as the atmospheric humidity is high. Nevertheless, if nothing is done, the thermal efficiency of the engine is unnecessarily reduced. If the thermal efficiency of the engine decreases, the fuel efficiency will decrease accordingly. Thus, in order not to reduce the heat efficiency unnecessarily when the atmospheric humidity is higher than that at the time of adaptation, it is necessary to increase the target compression ratio in the middle and high rotation speed range. This is because if the target compression ratio is increased, the thermal efficiency of the engine is improved.

  Therefore, in the fourth embodiment, a difference ΔH2 between the actual atmospheric humidity in the medium / high rotation speed region and the atmospheric humidity when the target compression ratio in the medium / high rotation speed region is adapted is calculated, and the medium / high rotation speed is calculated according to the difference ΔH2. Correct the target compression ratio in the speed range to the higher side. This improves fuel efficiency even when the actual atmospheric humidity is higher than the atmospheric humidity when the target compression ratio is set (adapted) in the medium to high rotational speed range during deceleration and reacceleration during constant speed driving on the highway. be able to.

  In the fourth embodiment, the target compression ratio in the medium / high rotation speed region is corrected based on the actual environmental condition parameters in the medium / high rotation speed region. As a result, even if the actual environmental condition parameters in the medium and high rotation speed range differ from the environmental condition parameters when the target compression ratio is set in the medium and high rotation speed range, deceleration reacceleration during constant speed driving on the highway Occasionally knocking can be suppressed and fuel consumption can be improved.

  In the fourth embodiment, the environmental condition parameter is at least one of intake air temperature, atmospheric temperature, engine cooling water temperature, engine lubricating oil temperature, atmospheric pressure, atmospheric humidity, and combustion chamber temperature. As a result, even when the actual combustion chamber temperature and atmospheric humidity in the medium and high rotation speed range differ from the combustion chamber temperature and atmospheric humidity when the target compression ratio is set in the medium and high rotation speed range, the highway constant speed Knocking can be suppressed and fuel consumption can be improved during deceleration and reacceleration during traveling.

  In the fourth embodiment, when the intercooler 61 is provided, the environmental condition parameter is the intercooler temperature. As a result, even if the actual intercooler temperature in the medium-high rotation speed range and the intercooler temperature when the target compression ratio is set in the medium-high rotation speed range are different, deceleration re-acceleration during constant speed driving on the highway Occasionally knocking can be suppressed and fuel consumption can be improved.

  In the embodiment, the case where the compression ratio variable mechanism is a multi-link type has been described, but this is not a limitation. The present invention is applicable not only to a turbocharger but also to a mechanical supercharger.

  The embodiment of the present invention has been described above. However, the above embodiment only shows a part of application examples of the present invention, and the technical scope of the present invention is limited to the specific configuration of the above embodiment. Absent.

1 Engine 51 Turbocharger (supercharger)
91 Engine controller (target compression ratio setting means, compression ratio control means)
93 Accelerator sensor (load detection means)
94 Crank angle sensor (rotational speed detection means)

Claims (7)

A turbocharger,
A variable compression ratio mechanism capable of changing the compression ratio of the engine;
A rotational speed detecting means for detecting the rotational speed of the engine;
Load detection means for detecting engine load;
Based on the detected engine speed and the detected engine load, target compression ratio setting means for setting the target compression ratio to a lower side than the low speed side area in the high speed side area even with the same load;
Compression ratio control means for controlling the compression ratio of the compression ratio variable mechanism so as to obtain the set target compression ratio;
Equipped with a,
The target compression ratio setting means is the time until the actual boost pressure reaches the set boost pressure or the actual engine at the time of deceleration reacceleration from the state where the vehicle is running at a constant speed in the region on the high rotation speed side. In the region on the high rotational speed side, the time until the torque reaches the set torque does not exceed the time until the compression ratio of the variable compression ratio mechanism reaches the target compression ratio after re-acceleration. Set the target compression ratio for
The compression ratio control means may obtain the set target compression ratio within a time until the actual supercharging pressure reaches the set supercharging pressure or within a time until the actual engine torque reaches the set torque. The engine control apparatus further controls a compression ratio of the compression ratio variable mechanism . When the total rotation speed area is divided into three areas, a low rotation speed area, a medium rotation speed area, and a high rotation speed area, the high rotation speed area is divided into the medium rotation speed area, the high rotation speed area, and the low rotation speed area. 2. The engine control apparatus according to claim 1, wherein the speed side region is a low rotation speed region. 3. The engine control device according to claim 2 , wherein the target compression ratio setting unit further sets the target compression ratio to a higher side as the rotational speed of the engine becomes higher in the high rotational speed range. 4. The target compression ratio setting means further sets the engine load for each target compression ratio so as not to exceed the operating point that becomes the compression ratio when the maximum combustion pressure is obtained, and this limit is not exceeded on the high rotation speed high load side. The engine control device according to claim 2 , wherein a curve representing a characteristic of the engine speed is set. The target compression ratio setting means corrects the target compression ratio in the medium rotation speed range and the high rotation speed range based on actual environmental condition parameters in the medium rotation speed range and the high rotation speed range. The engine control device according to any one of claims 2 to 4 . The environmental condition parameter is at least one of intake air temperature, atmospheric temperature, engine cooling water temperature, engine lubricating oil temperature, atmospheric pressure, atmospheric humidity, and combustion chamber temperature. 5. The engine control device according to 5. The engine control device according to claim 5 , wherein, when an intercooler is provided, the environmental condition parameter is an outlet temperature of the intercooler.

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