Homogeneous charge compression ignition internal combustion engine and control method thereof

Abstract

To obtain stable ignition and combustion by simple control, and a favorable exhaust composition.SOLUTION: An internal combustion engine which performs premixing compression self-ignition comprises a compression ratio variable mechanism which can change a mechanical compression ratio using a multi-link piston crank mechanism. The mechanical compression ratio is controlled so that the highest in-cylinder pressure crank angle θ Pmax approaches a target θ Pmax as an index indicating a combustion state. The target θ Pmax is calculated from an engine rotational speed (S1, S2) and load, and when a difference between the target θ Pmax and an actual θ Pmax is equal to or larger than a prescribed value, a compression ratio correction amount is acquired by multiplying a gain to the difference (S3, S5), and the target compression ratio is obtained by adding the compression ratio correction amount to a current actual compression ratio (S6). By the feedback control of the mechanical compression ratio, the θ Pmax is converged to the target θ Pmax.SELECTED DRAWING: Figure 2

Description

[Technical Field]

[0001] The present invention relates to a homogeneous charge compression ignition internal combustion engine that performs compression autoignition of a premixed air-fuel mixture, and a control method thereof. [Background technology]

[0002] For example, homogeneous charge compression ignition (HCCI) internal combustion engines have been proposed that use gasoline or other fuel as fuel and burn a premixed air-fuel mixture formed in a combustion chamber by compression ignition. Patent Document 1 discloses a homogeneous charge compression ignition (HCCI) internal combustion engine that controls the effective compression ratio by changing the intake valve closing timing in accordance with an estimated compression end temperature in order to achieve good HCCI. [Prior art documents] [Patent documents]

[0003] [Patent Document 1] Japanese Patent Application Publication No. 2019-39389 Summary of the Invention [Problem to be solved by the invention]

[0004] Changing the intake valve closing timing as described above changes not only the effective compression ratio but also the intake air amount and air-fuel ratio, making it impossible to appropriately control the combustion state through compression autoignition. If an attempt were made to correct the intake air amount and air-fuel ratio, which change with a change in the intake valve closing timing, by changing the fuel injection amount or throttle opening, the control would become complicated, making it difficult to ensure sufficient control responsiveness. [Means for solving the problem]

[0005] In this invention, A variable compression ratio mechanism is provided that can change the mechanical compression ratio. Detects the cylinder pressure, Actual combustion state due to compression autoignition of premixed air The maximum cylinder pressure crank angle is calculated as an index to show this. A target maximum cylinder pressure crank angle is calculated based on the engine rotation speed and the load. comparing the maximum cylinder pressure crank angle with the target maximum cylinder pressure crank angle, and correcting the target compression ratio of the compression ratio variable mechanism by a compression ratio correction amount calculated based on the difference between the maximum cylinder pressure crank angle and the target maximum cylinder pressure crank angle, so as to decrease the compression ratio if the maximum cylinder pressure crank angle is at a relatively early time, or to increase the compression ratio if the maximum cylinder pressure crank angle is at a relatively late time; According to this target compression ratio The mechanical compression ratio is controlled via the variable compression ratio mechanism. [Effects of the Invention]

[0006] According to this invention, the combustion state due to compression autoignition is brought closer to the target combustion state by controlling the mechanical compression ratio via a compression ratio variable mechanism, which simplifies control and makes it easy to obtain the target combustion state. [Brief explanation of the drawings]

[0007] [Figure 1] 1 is a diagram illustrating the configuration of an internal combustion engine according to an embodiment of the present invention; [Figure 2] 3 is a flowchart of compression ratio control. [Figure 3] 1 is a flowchart of intake air temperature control. [Figure 4] 1 is a timing chart of compression ratio control. [Figure 5] Time chart including intake air temperature control. [Figure 6] FIG. 1 is a functional block diagram of an embodiment. [Figure 7] 4 is a flowchart of injection amount control for each cylinder. [Figure 8] 4 is a time chart of an embodiment including injection amount control for each cylinder. [Figure 9] FIG. 2 is a functional block diagram of an injection amount control section for each cylinder. DETAILED DESCRIPTION OF THE INVENTION

[0008] An embodiment of the present invention will now be described in detail with reference to the accompanying drawings.

[0009] Fig. 1 shows the system configuration of an automotive internal combustion engine 1 to which the present invention is applied. This internal combustion engine 1 is a four-stroke cycle turbocharged internal combustion engine equipped with a compression ratio variable mechanism 2 that utilizes, for example, a multi-link piston-crank mechanism, and performs combustion by compression autoignition of a premixed air-fuel mixture as a homogeneous charge compression ignition internal combustion engine in a predetermined operating range. Note that in some operating ranges, combustion by spark ignition (including cases where compression autoignition is the main ignition but supplementary ignition is also performed) is performed.

[0010] A pair of intake valves 4 and a pair of exhaust valves 5 are arranged on the ceiling wall of the combustion chamber 3, and an ignition plug 6 for spark ignition in a specific operating range is arranged in the center surrounded by these intake valves 4 and exhaust valves 5. The combustion chamber 3 is provided with an in-cylinder pressure sensor 15 for detecting the in-cylinder pressure.

[0011] A fuel injection valve 8 that injects fuel toward the intake valve 4 is disposed for each cylinder in an intake port 7 that is opened and closed by the intake valve 4. The fuel injection valve 8 is an electromagnetic or piezoelectric injection valve that opens when a drive pulse signal is applied, and injects an amount of fuel that is substantially proportional to the pulse width of the drive pulse signal.

[0012] An electronically controlled throttle valve 11, whose opening is controlled by a control signal from an engine controller 10, is provided upstream of the collector section 9a of the intake passage 9 connected to the intake port 7. Further upstream of this is a compressor 12A of a turbocharger 12. An air flow meter 13 for detecting the intake air volume and an air cleaner 14 are provided upstream of the compressor 12A. A heat exchanger, i.e., a water-cooled intercooler 16, that cools the supercharged intake air by exchanging heat with the coolant, is built into the collector section 9a. An intercooler water temperature sensor 26 is provided in the intercooler 16 to detect the temperature of the coolant flowing through the intercooler 16. An intake air temperature sensor 27 is provided on the intake air outlet side of the intercooler 16 to detect the temperature of the intake air that has passed through the intercooler 16. In addition to the intercooler water temperature sensor 26, a coolant temperature sensor 28 is provided at an appropriate position, such as in the water jacket, of the internal combustion engine 1 to detect the temperature of the coolant flowing through the water jacket. The cooling water circuit flowing through the intercooler 16 may be part of the cooling water circuit flowing through the water jacket, or may be configured independent of each other.

[0013] The degree of heat exchange in the intercooler 16 can be variably adjusted via a cooling water valve, an intercooler cooling water pump, a cooling water bypass valve, or the like (not shown).

[0014] The turbine 12B of the turbocharger 12 disposed in the exhaust passage 21 is equipped with a wastegate valve 22. Catalytic devices 23 and 24 are disposed downstream of the turbine 12B in the exhaust passage 21, and a silencer 25 is disposed further downstream.

[0015] The compression ratio variable mechanism 2 utilizes a known multi-link piston crank mechanism, such as those disclosed in Japanese Patent Application Laid-Open Nos. 2005-127200 and 2004-116434, and is capable of changing the mechanical compression ratio by varying the top dead center position of the piston 30. The mechanical compression ratio is determined by the rotational position of a control shaft 31a moved by an electric actuator 31, and this rotational position is also detected as indicating the actual value of the mechanical compression ratio, i.e., the actual compression ratio. In other words, the electric actuator 31 also functions as an actual compression ratio sensor. In this compression ratio variable mechanism 2, the mechanical compression ratios of all cylinders are changed in the same way.

[0016] In addition to the above-mentioned in-cylinder pressure sensor 15, air flow meter 13, intercooler water temperature sensor 26, cooling water temperature sensor 28, intake air temperature sensor 27, and actual compression ratio sensor (electric actuator 31), detection signals from various sensors such as a crank angle sensor 32 for detecting engine speed, an accelerator opening sensor 33 for detecting the amount of depression of the accelerator pedal operated by the driver as a sensor for detecting torque demand by the driver, a boost pressure sensor 34 for detecting boost pressure, and an air-fuel ratio sensor 35 provided in the exhaust passage 21 are input to the engine controller 10. Based on these detection signals, the engine controller 10 optimally controls the amount of fuel injected by the fuel injector 8, the mechanical compression ratio by the compression ratio variable mechanism 2, the opening of the throttle valve 11, the opening of the wastegate valve 22, and the like, to achieve combustion by compression autoignition of the premixed mixture within a predetermined operating range.

[0017] In this embodiment, the maximum cylinder pressure crank angle (hereinafter abbreviated as θPmax), which is the crank angle at which the cylinder pressure reaches its maximum due to combustion, is used as an indicator of the combustion state due to compression ignition. Other indicators of the combustion state can also be used, such as the so-called MB50, which is the crank angle at which 50% of the fuel is burned, but θPmax can be most easily obtained using the cylinder pressure sensor 15. Then, the mechanical compression ratio is changed via the compression ratio variable mechanism 2 so that this θPmax approaches the target θPmax. By bringing the actual θPmax closer to the target θPmax in this way, stable ignition and combustion and a good exhaust gas composition can be achieved in homogeneous charge compression ignition combustion.

[0018] 2 shows a flowchart of the compression ratio control that is repeatedly executed by the engine controller 10. In step 1, the detection values ​​of various necessary sensors are read, and in step 2, the target θPmax is calculated based mainly on the engine rotation speed and load (required load). As described above, the target θPmax is a target θPmax that will result in stable ignition and combustion and a good exhaust composition.

[0019] Next, in step 3, the target θPmax is compared with the actual θPmax (for example, the θPmax detected in the immediately preceding cycle), and it is determined whether the difference (absolute value) between the two is equal to or greater than a predetermined value. If it is not equal to or greater than the predetermined value, the compression ratio correction amount is set to 0 (step 5). In other words, the predetermined value is a so-called dead zone. In the case of a multi-cylinder internal combustion engine, the actual θPmax may be represented by the value of one of the cylinders, but it is preferable to use the average of all the cylinders.

[0020] If the difference between the two is equal to or greater than a predetermined value, the process proceeds to step 4, where a compression ratio correction amount is calculated. For example, the compression ratio correction amount is calculated by multiplying an appropriate gain by the difference between the target θPmax and the actual θPmax. Here, if the actual θPmax is earlier than the target θPmax, the compression ratio correction amount is calculated to decrease the mechanical compression ratio, and conversely, if the actual θPmax is later than the target θPmax, the compression ratio correction amount is calculated to increase the mechanical compression ratio.

[0021] Then, the compression ratio correction amount is added to the current actual compression ratio to determine a target compression ratio (step 6), and the electric actuator 31 of the compression ratio variable mechanism 2 is driven in accordance with this target compression ratio (step 7).

[0022] In other words, in this embodiment, the mechanical compression ratio is feedback-controlled based on the deviation between the target θPmax and the actual θPmax. The time chart in Figure 4 shows the changes in θPmax and the mechanical compression ratio obtained by such feedback control, and the mechanical compression ratio repeatedly fluctuates between high and low levels so that the actual θPmax converges to the target θPmax. This maintains stable homogeneous charge compression ignition combustion.

[0023] Figure 3 is a flowchart of intake air temperature control that is executed in parallel with the compression ratio control of Figure 2. This controls the temperature of the intake air taken into the combustion chamber 3 so that the mechanical compression ratio, which changes as shown in Figure 4, approaches an appropriate reference mechanical compression ratio.

[0024] The flowchart of Fig. 3 is repeatedly executed by the engine controller 10. In step 11, detected values ​​of various necessary sensors are read, and in step 12, a reference mechanical compression ratio is calculated mainly based on the engine speed and load (required load). The reference mechanical compression ratio is a reference intermediate value that is set so as to leave an appropriate controllable margin on both the high compression ratio side and the low compression ratio side within the compression ratio variable range that can be changed by the compression ratio variable mechanism 2, and is optimally set based on the engine speed and load, taking into account changes in the compression ratio during transients, etc. Alternatively, the reference mechanical compression ratio may be set to a constant value for simplicity.

[0025] Next, in step 13, the reference mechanical compression ratio is compared with the actual compression ratio (the actual compression ratio as a result of feedback control as shown in FIG. 4), and it is determined whether the difference (absolute value) between the two is equal to or greater than a predetermined value. If it is not equal to or greater than the predetermined value, the intake air temperature correction amount is set to 0 (step 15). In other words, the predetermined value is a so-called dead band.

[0026] If the difference between the two is equal to or greater than a predetermined value, the process proceeds to step 14, where an intake air temperature correction amount is calculated. For example, the intake air temperature correction amount is calculated by multiplying the difference between the reference mechanical compression ratio and the actual compression ratio by an appropriate gain. Here, the intake air temperature correction amount is calculated so that the intake air temperature increases if the actual compression ratio is higher than the reference mechanical compression ratio, and conversely, so that the intake air temperature decreases if the actual compression ratio is lower than the reference mechanical compression ratio.

[0027] Then, the intake air temperature correction amount is added to the current intake air temperature (the value detected by the intake air temperature sensor 27) to determine a target intake air temperature (step 16), and the heat exchange amount of the water-cooled intercooler 16 is adjusted in accordance with this target intake air temperature (step 17). For example, the heat exchange amount is increased or decreased via a cooling water valve, an intercooler cooling water pump, a cooling water bypass valve, or the like (not shown) to bring the intake air temperature closer to the target intake air temperature. Note that the specific control of the intake air temperature is not a main part of the present invention, and the intake air temperature may be adjusted by any known appropriate method or means.

[0028] The intake air temperature may be controlled by variably controlling the temperature of the cooling water flowing through the water jacket of the internal combustion engine 1. It is also possible to increase the intake air temperature by using an appropriate heating means.

[0029] That is, in this embodiment, the intake air temperature is feedback-controlled based on the deviation between the reference mechanical compression ratio and the actual compression ratio. The time chart of FIG. 5 shows the changes in θPmax, the changes in the mechanical compression ratio, and the changes in the intake air temperature (the outlet temperature of the intercooler 16) obtained by such feedback control. Note that this time chart has a compressed time axis compared to the time chart of FIG. 4. The initial portion of the time chart of FIG. 5 (i.e., the portion immediately after the start of control) is the same as FIG. 4. As shown in FIG. 4, as a result of feedback control of the mechanical compression ratio, the actual θPmax attempts to converge to the target θPmax. However, at this time, the mechanical compression ratio may attempt to stabilize at a point deviating from the reference mechanical compression ratio. When the actual compression ratio deviates from the reference mechanical compression ratio as a result of feedback control, applying the intake air temperature control shown in FIG. 3 changes the temperature of the intake air flowing into the cylinder, and as a result, the actual compression ratio approaches the reference mechanical compression ratio.

[0030] In the example of FIG. 5, immediately after the start of control, the actual compression ratio tends to be lower than the reference mechanical compression ratio. In response to this, the intake air temperature is controlled to decrease based on the deviation between the actual compression ratio and the reference mechanical compression ratio. When the intake air temperature decreases in this way, the pressure increase during homogeneous charge compression ignition slows down, and in order to bring the actual θPmax closer to the target θPmax, the mechanical compression ratio is controlled to become relatively higher. Therefore, as shown in the example of FIG. 5, the actual compression ratio gradually approaches the reference mechanical compression ratio.

[0031] Note that the change in intake air temperature, and therefore the change in the actual compression ratio based on the intake air temperature, is slow. Therefore, the actual compression ratio repeatedly rises and falls due to feedback control based on the deviation between the target θPmax and the actual θPmax, and gradually approaches the reference mechanical compression ratio over a relatively long period of time.

[0032] By keeping the control center of the compression ratio control near the reference mechanical compression ratio in this way, it is possible to leave an appropriate controllable margin on both the low compression ratio side and the high compression ratio side, and it is possible to respond appropriately to cases where the deviation of θPmax becomes large, for example, during transient conditions.

[0033] FIG. 6 is a functional block diagram illustrating the compression ratio control including the intake air temperature control described above. As shown in the figure, the engine controller 10 functionally includes a target θPmax calculation unit 41, a reference compression ratio calculation unit 42, a compression ratio correction amount calculation unit 43, a target compression ratio calculation unit 44, an intake air temperature correction amount calculation unit 45, and a target intake air temperature calculation unit 46. The target θPmax calculation unit 41 and the reference compression ratio calculation unit 42 output the target θPmax and the reference mechanical compression ratio, respectively, based on the engine speed detected by the crank angle sensor 32 and the required load calculated from the output value of the accelerator opening sensor 33. The compression ratio correction amount calculation unit 43 outputs the required compression ratio correction amount based on the difference between the target θPmax and the actual θPmax (e.g., the average of all cylinders). The target compression ratio calculation unit 44 adds the compression ratio correction amount to the actual compression ratio at that time to output the target compression ratio. The actuator 31 of the compression ratio variable mechanism 2 is driven in accordance with this target compression ratio. The actuator 31 is feedback controlled in an appropriate manner to achieve the target compression ratio.

[0034] Furthermore, an intake air temperature correction amount calculation unit 45 compares the reference mechanical compression ratio with the actual compression ratio and outputs the necessary intake air temperature correction amount based on the difference between the two so that the actual compression ratio approaches the reference mechanical compression ratio. A target intake air temperature calculation unit 46 adds the intake air temperature correction amount to the current intake air temperature and outputs a target intake air temperature. An appropriate intake air temperature control device 47 is driven in accordance with this target intake air temperature. As described above, the intake air temperature control device 47 is configured by the intercooler 16, a cooling device for the internal combustion engine 1, or another appropriate mechanism.

[0035] Next, a second embodiment will be described in which, in addition to the above-described compression ratio control, variations in the combustion state of each cylinder (e.g., differences in θPmax between cylinders) are reduced in a multi-cylinder internal combustion engine. In the above-described variable compression ratio mechanism 2, the compression ratios of all cylinders are changed in the same way, so it is not possible to fine-tune the combustion state of each cylinder (e.g., θPmax or MB50) by adjusting the compression ratio. Therefore, in this second embodiment, in addition to the compression ratio control for all cylinders, the combustion state of each cylinder is optimized by correcting the fuel injection amount for each cylinder.

[0036] FIG. 7 is a flowchart of injection amount control repeatedly executed by the engine controller 10 in parallel with the compression ratio control of FIG. 2. It is also possible to combine this with the compression ratio control including the intake air temperature control shown in FIG. 3. In step 21, the detection values ​​of various necessary sensors are read, and in step 22, the reference fuel injection amount is calculated based mainly on the engine speed and load (required load). Furthermore, in step 22, the cylinder average θPmax, which is the average of all cylinders of actual θPmax, is calculated. It is to be noted that the processing shown in the flowchart of FIG. 7 is executed for each cylinder in accordance with the combustion order, and the cylinder average θPmax can be calculated, for example, as a moving average. "N" represents the cylinder number.

[0037] Next, in step 23, the cylinder average θPmax is compared with the actual θPmax of cylinder #N, and it is determined whether the difference (absolute value) between the two is equal to or greater than a predetermined value. If it is not equal to or greater than the predetermined value, the fuel injection amount correction amount for cylinder #N is set to 0 (step 25). In other words, the predetermined value is a so-called dead band.

[0038] If the difference between the two is equal to or greater than a predetermined value, the process proceeds to step 24, where the fuel injection amount correction amount for cylinder #N is calculated. For example, the compression ratio correction amount for cylinder #N is determined by multiplying the difference between the cylinder average θPmax and the actual θPmax of cylinder #N by an appropriate gain. Here, the fuel injection amount correction amount is calculated to decrease the fuel injection amount if the actual θPmax of cylinder #N is earlier than the cylinder average θPmax, and conversely, to increase the fuel injection amount if the actual θPmax of cylinder #N is later than the cylinder average θPmax.

[0039] Then, the fuel injection amount correction amount is added to the current fuel injection amount for cylinder #N (the initial fuel injection amount is the reference fuel injection amount) to determine the fuel injection amount for cylinder #N (step 26), and the fuel injection valve 8 for cylinder #N is driven in accordance with this fuel injection amount (step 27). Note that actual injection is performed at a predetermined injection timing.

[0040] That is, in this embodiment, the mechanical compression ratio of all cylinders is controlled based on the cylinder average θPmax, while the fuel injection amount is corrected based on the actual θPmax of each cylinder so that the difference in θPmax between cylinders becomes small.

[0041] The time chart in Figure 8 shows the changes in θPmax, mechanical compression ratio, and fuel injection amount obtained by the control of the second embodiment. For the sake of explanation, the cylinder average θPmax etc. are shown assuming there are two cylinders (cylinder #1 and cylinder #2).

[0042] The θPmax column in Figure 8 shows the θPmax of cylinder #1, the θPmax of cylinder #2, and the average of the two, the cylinder-averaged θPmax. The mechanical compression ratio is feedback-controlled so that the cylinder-averaged θPmax converges to the target θPmax. Meanwhile, the fuel injection amount of each cylinder is feedback-controlled so that the deviation between the θPmax of each cylinder and the cylinder-averaged θPmax is reduced. In the illustrated example, the θPmax of cylinder #1 tends to be more advanced than the θPmax of cylinder #2, so the fuel injection amount of cylinder #1 is corrected downward, and the fuel injection amount of cylinder #2 is corrected upward. As a result, the θPmax of cylinder #1 and the θPmax of cylinder #2 approach each other with some overshoot and simultaneously converge to the target θPmax. Therefore, each cylinder maintains an appropriate combustion state (θPmax in this example).

[0043] FIG. 9 is a functional block diagram showing the injection amount control portion described above. As shown in the figure, the engine controller 10 functionally includes a base fuel injection amount calculation unit 51, a cylinder average θPmax calculation unit 52, cylinder-specific fuel injection amount correction amount calculation units 53 (53-1 to 53-n), and cylinder-specific fuel injection amount calculation units 54 (54-1 to 54-n). The base fuel injection amount calculation unit 51 outputs a base fuel injection amount based on the engine speed detected by the crank angle sensor 32 and the required load calculated from the output value of the accelerator opening sensor 33. Note that other correction factors may be added in addition to the engine speed and load. The cylinder average θPmax calculation unit 52 calculates the average of all cylinders of the actual θPmax of each cylinder calculated based on the cylinder pressure sensor 15 for each cylinder, i.e., the cylinder average θPmax.

[0044] A fuel injection amount correction amount calculation unit 53 (53-1 to 53-n) for each cylinder outputs the required fuel injection amount correction amount for each cylinder based on the difference between the cylinder average θPmax and the actual θPmax for that cylinder. A fuel injection amount calculation unit 54 (54-1 to 54-n) for each cylinder adds the fuel injection amount correction amount to the fuel injection amount for that cylinder (the reference fuel injection amount for the first injection) and outputs the fuel injection amount for each cylinder. The fuel injection valve 8 for each cylinder is driven according to this fuel injection amount.

[0045] Although one embodiment of the present invention has been described above in detail, the present invention is not limited to the above embodiment and various modifications are possible. For example, in the above example, θPmax is used as an index showing the combustion state of homogeneous charge compression ignition, but other indexes such as MB50 may also be used. Furthermore, in the above example, control of the mechanical compression ratio and the like is described in the form of feedback control using relatively simple proportional control, but the present invention is not limited to this and any known appropriate control method may be used. [Explanation of symbols]

[0046] 1...Internal combustion engine 2...Variable compression ratio mechanism 8...Fuel injection valve 10...Engine controller 15...Cylinder pressure sensor 16...Water-cooled intercooler 27...Intake air temperature sensor 32...Crank angle sensor 33...Accelerator opening sensor

Claims

1. A variable compression ratio mechanism is provided that can change the mechanical compression ratio. The cylinder pressure is detected, and the maximum cylinder pressure crank angle is calculated as an index showing the actual combustion state due to compression autoignition of the premixed air-fuel mixture. A target maximum cylinder pressure crank angle is calculated based on the engine rotation speed and the load. comparing the maximum cylinder pressure crank angle with the target maximum cylinder pressure crank angle, and correcting the target compression ratio of the compression ratio variable mechanism by a compression ratio correction amount calculated based on the difference between the maximum cylinder pressure crank angle and the target maximum cylinder pressure crank angle, so as to decrease the compression ratio if the maximum cylinder pressure crank angle is at a relatively early time, or to increase the compression ratio if the maximum cylinder pressure crank angle is at a relatively late time; controlling the mechanical compression ratio via the compression ratio variable mechanism in accordance with the target compression ratio; A method for controlling a homogeneous charge compression ignition internal combustion engine.

2. The variable compression ratio mechanism is a mechanism that changes the mechanical compression ratio of all cylinders in the same way, an average of the maximum cylinder pressure crank angles of all cylinders is used as the maximum cylinder pressure crank angle to be compared with the target maximum cylinder pressure crank angle; 2. A method for controlling a homogeneous charge compression ignition internal combustion engine according to claim 1.

3. A variable compression ratio mechanism is provided that can change the mechanical compression ratio, Detects the actual combustion state due to compression autoignition of premixed air-fuel mixture, controlling the mechanical compression ratio via the compression ratio variable mechanism so that the detected combustion state approaches a target combustion state; A control method for a homogeneous charge compression ignition internal combustion engine, comprising: A reference mechanical compression ratio is set within a variable compression ratio range, controlling the temperature of intake air taken into the combustion chamber so that the mechanical compression ratio controlled based on the combustion state approaches a reference mechanical compression ratio; A method for controlling a homogeneous charge compression ignition internal combustion engine.

4. The temperature of the intake air is controlled using a heat exchanger provided in the intake passage.

4. The method for controlling a homogeneous charge compression ignition internal combustion engine according to claim 3.

5. Controlling the intake air temperature through variable control of the engine coolant temperature; 4. The method for controlling a homogeneous charge compression ignition internal combustion engine according to claim 3.

6. A variable compression ratio mechanism is provided that can change the mechanical compression ratio, Detects the actual combustion state due to compression autoignition of premixed air-fuel mixture, controlling the mechanical compression ratio via the compression ratio variable mechanism so that the detected combustion state approaches a target combustion state; A control method for a homogeneous charge compression ignition internal combustion engine, comprising: In a multi-cylinder internal combustion engine, the mechanical compression ratio of all cylinders is controlled based on the average combustion state of the cylinders; Correcting the fuel injection amount for each cylinder based on the difference in combustion state between the cylinders so as to reduce the difference. A method for controlling a homogeneous charge compression ignition internal combustion engine.

7. an internal combustion engine that performs compression autoignition of a premixed air-fuel mixture in at least a part of an operating range; a variable compression ratio mechanism capable of changing the mechanical compression ratio; A controller; Equipped with The above controller is The cylinder pressure is detected, and the maximum cylinder pressure crank angle is calculated as an index showing the actual combustion state due to compression autoignition of the premixed air-fuel mixture. A target maximum cylinder pressure crank angle is calculated based on the engine rotation speed and the load. comparing the maximum cylinder pressure crank angle with the target maximum cylinder pressure crank angle, and correcting the target compression ratio of the compression ratio variable mechanism by a compression ratio correction amount calculated based on the difference between the maximum cylinder pressure crank angle and the target maximum cylinder pressure crank angle, so as to decrease the compression ratio if the maximum cylinder pressure crank angle is at a relatively early time, or to increase the compression ratio if the maximum cylinder pressure crank angle is at a relatively late time; controlling the mechanical compression ratio via the compression ratio variable mechanism in accordance with the target compression ratio; Homogeneous charge compression ignition internal combustion engine.

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