JP2019039382A - Signal processing device for engine - Google Patents

Abstract

To provide a signal processing device for an engine, capable of more properly correcting a cylinder pressure.SOLUTION: The signal processing device for the engine includes cylinder pressure detection means SN2 for detecting a cylinder pressure as a pressure in a cylinder 3, and intake pressure detection means SN6 for detecting an intake pressure as a pressure in an intake passage 30. Correction means 122 corrects the cylinder pressure detected by the cylinder pressure detection means SN2 on the basis of the intake pressure detected by the intake pressure detection means SN6 during a reference period set near an intake bottom dead center.SELECTED DRAWING: Figure 9

Description

  The present invention relates to a signal processing device provided in an engine including an engine body in which a cylinder is formed and an intake passage through which intake air introduced into the cylinder flows.

  Conventionally, it has been studied to provide an in-cylinder pressure detecting means for detecting the pressure in the cylinder in the engine body, and to use the in-cylinder pressure detected by the in-cylinder pressure detecting means for controlling each part of the engine body. For example, the ignition timing and the like are controlled by calculating the heat generation rate and heat generation amount in the cylinder using the in-cylinder pressure.

  Here, the absolute value of the in-cylinder pressure may be required to calculate the heat generation rate and the like. Therefore, if the pressure detected by the in-cylinder pressure detection means is not an absolute value (in the case of a signal proportional to the pressure, etc.) or if the detected pressure is likely to deviate from the actual pressure, it will be detected. It is necessary to correct the measured pressure.

  On the other hand, in Patent Document 1, the specific heat ratio of the gas in the cylinder, the in-cylinder volume at two different crank angles and the in-cylinder pressure detected by the in-cylinder pressure detecting means, and the specific heat ratio of the gas in the cylinder Is applied to the Poisson's relational expression to correct the in-cylinder pressure detected by the in-cylinder pressure detecting means.

JP 2014-141931 A

  In the apparatus of Patent Document 1, the correction of the in-cylinder pressure includes an error in the specific heat ratio of the gas in the cylinder and an in-cylinder volume error. Therefore, there is a problem that the correction accuracy is not sufficiently high.

  The present invention has been made in view of the above circumstances, and an object thereof is to provide an engine signal processing device capable of more appropriately correcting the in-cylinder pressure.

  In order to solve the above-mentioned problem, the present invention is a signal processing device provided in an engine including an engine body in which a cylinder is formed and an intake passage through which intake air introduced into the cylinder flows. An in-cylinder pressure detecting means for detecting an in-cylinder pressure that is an internal pressure; an intake pressure detecting means for detecting an intake pressure that is a pressure in the intake passage; and the intake pressure during a reference period set near the intake bottom dead center. Provided is a signal processing apparatus for an engine comprising correction means for correcting the in-cylinder pressure detected by the in-cylinder pressure detection means based on the intake pressure detected by the atmospheric pressure detection means.

  According to this device, the in-cylinder pressure can be accurately obtained by correcting the in-cylinder pressure more appropriately with a simple configuration in which the intake pressure detected by the intake pressure detection means is used. Specifically, in the reference period set near the intake bottom dead center, the pressure in the intake passage and the pressure in the cylinder, that is, the cylinder pressure are substantially the same. Therefore, by correcting the in-cylinder pressure detected by the in-cylinder pressure detecting means using the pressure in the intake passage during the reference period, the detected in-cylinder pressure can be easily corrected to a value closer to the actual value. .

  In the above configuration, the engine body includes a plurality of the cylinders, the intake passage communicates with the cylinders, and the correction unit detects a plurality of detection points detected by the intake pressure detection unit during the reference period. It is preferable to calculate an average value of the intake pressure and correct the in-cylinder pressure detected by the in-cylinder pressure detecting means using the average value.

  According to this configuration, the in-cylinder pressure detected by the in-cylinder pressure detecting unit can be accurately corrected even in a multi-cylinder engine in which the intake passage communicates with a plurality of cylinders. In other words, when the intake passage communicates with a plurality of cylinders, the pressure in the intake passage, that is, the intake pressure may pulsate, and there may be a deviation between the in-cylinder pressure and the intake pressure. For example, since the in-cylinder pressure is corrected using the average value of the intake pressure, the in-cylinder pressure can be corrected more appropriately while suppressing the deviation.

  In the above configuration, it is preferable that the reference period is set to a period earlier than the closing timing of the intake valve that opens and closes the intake port communicating with the cylinder.

  According to this configuration, the in-cylinder pressure detected by using the intake pressure at the time when the difference between the intake pressure and the in-cylinder pressure is more reliably reduced before the intake valve is closed is corrected. This correction can be performed more appropriately.

  In the above configuration, the correction means may set the reference period to a period earlier than the intake bottom dead center when the closing timing of the intake valve is retarded from the intake bottom dead center. Preferred (claim 4).

  In this configuration, when the intake valve closes on the retard side from the intake bottom dead center and the intake air is blown back into the intake passage from the inside of the cylinder, the intake pressure before the intake bottom dead center is The in-cylinder pressure detected by the in-cylinder pressure detecting means is corrected using the intake pressure during a period in which the intake air does not blow back into the intake passage. Therefore, it is possible to prevent the correction from being performed by the intake pressure in a state where the deviation from the in-cylinder pressure is large due to the return of the intake air, and the in-cylinder pressure can be corrected more appropriately.

  As described above, since the in-cylinder pressure can be obtained with higher accuracy in the present invention, the amount of heat generated in the cylinder and the combustion center-of-gravity timing can be accurately estimated using this in-cylinder pressure. Therefore, the present invention estimates the amount of heat generated in the cylinder based on the in-cylinder pressure after being corrected by the correcting means, and burns 50% of the fuel supplied to the cylinder in one combustion cycle. It is effective if it is applied to a device provided with combustion gravity center timing estimating means for estimating the combustion gravity center timing at the time of the above (claim 5).

  As described above, according to the engine signal processing apparatus of the present invention, the in-cylinder pressure can be corrected more appropriately with a simple configuration.

1 is a schematic configuration diagram of an engine system to which an engine signal processing device according to an embodiment of the present invention is applied. It is a schematic block diagram around an engine body. It is the figure which showed together sectional drawing of an engine main body, and the top view of a piston. It is a block diagram which shows the control system of an engine. It is the figure which showed the operating area | region of the engine. It is the figure which illustrated fuel injection timing, ignition timing, and a combustion waveform in each operation field. It is a graph which shows the waveform of the heat release rate at the time of SPCCI combustion (partial compression ignition combustion). It is a figure for demonstrating the signal processing of a cylinder pressure sensor. It is the flowchart which showed the procedure of the conversion process. It is the figure which showed the waveform of intake pressure and in-cylinder pressure.

  Hereinafter, an engine signal processing device according to an embodiment of the present invention will be described with reference to the drawings.

(1) Overall Configuration of Engine FIGS. 1 to 3 are views showing a preferred embodiment of an engine to which the signal processing device of the present invention is applied. The engine shown in this figure is a four-cycle gasoline direct injection engine mounted on a vehicle as a driving power source. The engine body 1, an intake passage 30 through which intake air introduced into the engine body 1 circulates, An exhaust passage 40 through which exhaust gas discharged from the engine body 1 flows and an EGR device 50 that recirculates part of the exhaust gas flowing through the exhaust passage 40 to the intake passage 30 are provided.

  The engine body 1 includes a cylinder block 3 in which a cylinder 2 is formed, a cylinder head 4 attached to the upper surface of the cylinder block 3 so as to close the cylinder 2 from above, and a reciprocating slide on the cylinder 2. And an inserted piston 5. As shown in FIG. 2, the engine main body 1 is of a multi-cylinder type, and has four cylinders arranged in a row in the illustrated example.

  A combustion chamber 6 is defined above the piston 5, and fuel mainly composed of gasoline is supplied to the combustion chamber 6 by injection from an injector 15 described later. The supplied fuel burns while being mixed with air in the combustion chamber 6, and the piston 5 pushed down by the expansion force due to the combustion reciprocates in the vertical direction. In addition, the fuel injected into the combustion chamber 6 should just contain gasoline as a main component, for example, in addition to gasoline, it may contain subcomponents, such as bioethanol.

  Below the piston 5, a crankshaft 7 that is an output shaft of the engine body 1 is provided. The crankshaft 7 is connected to the piston 5 via a connecting rod 8 and is rotationally driven around the central axis according to the reciprocating motion (vertical motion) of the piston 5.

  The geometric compression ratio of the cylinder 2, that is, the ratio between the volume of the combustion chamber 6 when the piston 5 is at the top dead center and the volume of the combustion chamber when the piston 5 is at the bottom dead center is the SPCCI combustion (described later) A value suitable for partial compression ignition combustion) is set to 13 or more and 30 or less. More specifically, the geometric compression ratio of the cylinder 2 is set to 14 or more and 17 or less in the case of a regular specification using a gasoline fuel having an octane number of about 91, and a high-octane specification using a gasoline fuel having an octane number of about 96. In some cases, it is preferably set to 15 or more and 18 or less.

  The cylinder block 3 is provided with a crank angle sensor SN1 that detects the rotation angle (crank angle) of the crankshaft 7 and the rotation speed (engine rotation speed) of the crankshaft 7.

  The cylinder head 4 is provided with an intake port 9 and an exhaust port 10 that open to the combustion chamber 6, an intake valve 11 that opens and closes the intake port 9, and an exhaust valve 12 that opens and closes the exhaust port 10.

  As shown in FIGS. 2 and 3, the valve type of the engine of this embodiment is a four-valve type of 2 intake valves × 2 exhaust valves. That is, the intake port 9 has a first intake port 9A and a second intake port 9B, and the exhaust port 10 has a first exhaust port 10A and a second exhaust port 10B. The intake valve 11 is provided so as to open and close the first intake port 9A and the second intake port 9B (two in total), and the exhaust valve 12 opens and closes the first exhaust port 10A and the second exhaust port 10B, respectively. (2 in total).

  The second intake port 9B is provided with a swirl valve 18 that can be opened and closed. The swirl valve 18 opens and closes to increase the strength of the swirl flow in the cylinder 2 (the swirl flow swirling around the axis of the cylinder 2). It can be changed.

  The intake valve 11 and the exhaust valve 12 are driven to open and close in conjunction with the rotation of the crankshaft 7 by valve mechanisms 13 and 14 including a pair of camshafts and the like disposed in the cylinder head 4.

  The valve operating mechanism 13 for the intake valve 11 incorporates an intake VVT 13a that can change at least the opening timing of the intake valve 11. Similarly, the valve mechanism 14 for the exhaust valve 12 incorporates an exhaust VVT 14a that can change at least the closing timing of the exhaust valve 12. By controlling the intake VVT 13a and the exhaust VVT 14a, in this embodiment, it is possible to adjust the valve overlap period in which both the intake valve 11 and the exhaust valve 12 are opened across the exhaust top dead center. By adjusting the valve overlap period, the amount of burned gas (internal EGR gas) remaining in the combustion chamber 6 can be adjusted. The intake VVT 13a (exhaust VVT 14a) may be a variable mechanism that changes only the closing timing (opening timing) while fixing the opening timing (closing timing) of the intake valve 11 (exhaust valve 12). A phase-type variable mechanism that simultaneously changes the opening timing and closing timing of the intake valve 11 (exhaust valve 12) may be used.

  As shown in FIG. 3, a cavity 20 is formed on the crown surface of the piston 5, in which a relatively wide region including the central portion is recessed on the opposite side (downward) from the cylinder head 4. A substantially conical bulge 20a is formed at the center of the cavity 20 so as to bulge relatively upward, and both sides in the radial direction sandwiching the bulge 20a are recesses each having a bowl-like cross section. . In other words, the cavity 20 is a donut-shaped concave portion in plan view formed so as to surround the raised portion 20a. A region on the outer side in the radial direction from the cavity 20 in the crown surface of the piston 5 is a squish portion 21 formed of an annular flat surface.

  The cylinder head 4 ignites a mixture of an injector 15 that injects fuel (mainly gasoline) into the combustion chamber 6, and fuel that is injected from the injector 15 into the combustion chamber 6 and air that is introduced into the combustion chamber 6. A spark plug 16 is provided.

  The injector 15 is a multi-hole type injector having a plurality of injection holes at its tip, and can inject fuel radially from the plurality of injection holes (F in FIG. Represents the spray of fuel injected from the hole). The injector 15 is provided so that the tip thereof faces the central portion (the raised portion 20 a) of the crown surface of the piston 5.

  The spark plug 16 is disposed at a position somewhat shifted to the intake side with respect to the injector 15. The tip (electrode part) of the spark plug 16 is set at a position overlapping the cavity 20 in plan view.

  As shown in FIG. 1, the cylinder head 4 is further provided with an in-cylinder pressure sensor (in-cylinder pressure detecting means) SN2 for detecting an in-cylinder pressure that is a pressure in the combustion chamber 6 (cylinder 2). In the present embodiment, one in-cylinder pressure sensor SN2 is provided for each cylinder 2.

  The in-cylinder pressure sensor SN2 is attached to the cylinder head 4 so as to face the inside of the combustion chamber 6 from the ceiling surface of the combustion chamber 6. Further, the in-cylinder pressure sensor SN <b> 2 is arranged so that the tip thereof faces the vicinity of the center of the crown surface of the piston 5.

  The in-cylinder pressure sensor SN2 outputs a voltage corresponding to the pressure in the combustion chamber 6. For example, the in-cylinder pressure sensor SN2 has a passage communicating with the combustion chamber 6 formed at the tip thereof, a diaphragm provided at the end of the passage and displaced in accordance with the pressure in the passage, and attached to the diaphragm. And a piezoelectric element. In such an in-cylinder pressure sensor SN2, the diaphragm is displaced according to the pressure in the combustion chamber 6 and the passage communicating therewith, and the cylinder corresponding to the voltage corresponding to the displacement amount of the diaphragm, that is, the pressure in the combustion chamber 6. The piezoelectric element outputs a voltage corresponding to the internal pressure.

  Thus, the in-cylinder pressure sensor SN2 outputs a voltage, and the voltage output from the in-cylinder pressure sensor SN2 is converted into a pressure by a pressure conversion unit (correction means) 102 of the ECU 100 described later.

  As shown in FIG. 1, the intake passage 30 is connected to one side surface of the cylinder head 4 so as to communicate with the intake port 9. Air (fresh air) taken from the upstream end of the intake passage 30 is introduced into the combustion chamber 6 through the intake passage 30 and the intake port 9.

  In the intake passage 30, in order from the upstream side, an air cleaner 31 that removes foreign matter in the intake air, an openable / closable throttle valve 32 that adjusts the flow rate of intake air, a supercharger 33 that supercharges intake air, and intake air An intercooler 35 for cooling and a surge tank 36 are provided. As shown in FIG. 2, the intake passage 30 is branched into four passages downstream from the surge tank 36, and each branch passage communicates with the two intake ports 9 of each cylinder 2.

  Each part of the intake passage 30 is provided with an air flow sensor SN3 for detecting the flow rate of intake air and first and second intake air temperature sensors SN4 and SN5 for detecting the temperature of intake air. The air flow sensor SN3 and the first intake air temperature sensor SN4 are provided in a portion of the intake passage 30 between the air cleaner 31 and the throttle valve 32, and detect the flow rate and temperature of the intake air passing through the portion. The second intake air temperature sensor SN5 is provided in a portion of the intake passage 30 between the supercharger 33 and the intercooler 35, and detects the temperature of intake air that passes through the portion.

  The intake passage 30 is provided with an intake pressure sensor (intake pressure detection means) SN6 that detects an intake pressure that is the pressure in the intake passage and is the pressure of the intake air introduced into the cylinder 2. The intake pressure sensor SN6 is provided in the surge tank 36 and detects the pressure of intake air in the surge tank 36.

  The supercharger 33 is a mechanical supercharger (supercharger) mechanically linked to the engine body 1. Although the specific form of the supercharger 33 is not particularly limited, for example, any of the known superchargers such as a Rishorum type, a roots type, or a centrifugal type can be used as the supercharger 33.

  Between the supercharger 33 and the engine main body 1, an electromagnetic clutch 34 capable of electrically switching between fastening and releasing is interposed. When the electromagnetic clutch 34 is engaged, driving force is transmitted from the engine body 1 to the supercharger 33, and supercharging by the supercharger 33 is performed. On the other hand, when the electromagnetic clutch 34 is released, the transmission of the driving force is interrupted and the supercharging by the supercharger 33 is stopped.

  The intake passage 30 is provided with a bypass passage 38 for bypassing the supercharger 33. The bypass passage 38 connects the surge tank 36 and an EGR passage 51 described later to each other. A bypass valve 39 that can be opened and closed is provided in the bypass passage 38.

  The exhaust passage 40 is connected to the other side of the cylinder head 4 so as to communicate with the exhaust port 10. Burned gas (exhaust gas) generated in the combustion chamber 6 is discharged to the outside through the exhaust port 10 and the exhaust passage 40.

  A catalytic converter 41 is provided in the exhaust passage 40. The catalytic converter 41 includes a three-way catalyst 41a for purifying harmful components (HC, CO, NOx) contained in the exhaust gas flowing through the exhaust passage 40, and particulate matter (PM) contained in the exhaust gas. And a GPF (gasoline particulate filter) 31b for collecting gas. Note that another catalytic converter incorporating an appropriate catalyst such as a three-way catalyst or a NOx catalyst may be added downstream of the catalytic converter 41.

  The EGR device 50 includes an EGR passage 51 that connects the exhaust passage 40 and the intake passage 30, and an EGR cooler 52 and an EGR valve 53 that are provided in the EGR passage 51. The EGR passage 51 connects a portion of the exhaust passage 40 downstream of the catalytic converter 41 and a portion of the intake passage 30 between the throttle valve 32 and the supercharger 33. The EGR cooler 52 cools the exhaust gas (external EGR gas) recirculated from the exhaust passage 40 to the intake passage 30 through the EGR passage 51 by heat exchange. The EGR valve 53 is provided so as to be openable and closable in the EGR passage 51 on the downstream side (closer to the intake passage 30) than the EGR cooler 52, and adjusts the flow rate of the exhaust gas flowing through the EGR passage 51.

  The EGR passage 51 is provided with a differential pressure sensor SN7 for detecting the difference between the upstream pressure and the downstream pressure of the EGR valve 53.

(2) Control System FIG. 4 is a block diagram showing an engine control system. The ECU 100 shown in the figure includes an I / F circuit 110 (see FIG. 8) to which detection signals of various sensors are input, and a microcomputer 120 (see FIG. 8) for comprehensively controlling the engine.

  The ECU 100 receives detection signals from various sensors. For example, the ECU 100 is electrically connected to the crank angle sensor SN1, the in-cylinder pressure sensor SN2, the airflow sensor SN3, the first and second intake air temperature sensors SN4 and SN5, the intake pressure sensor SN6, and the differential pressure sensor SN7. Information detected by these sensors is sequentially input to the ECU 100. Further, the vehicle is provided with an accelerator sensor SN8 that detects the opening of an accelerator pedal (not shown) operated by a driver who drives the vehicle, and a detection signal from the accelerator sensor SN8 is also input to the ECU 100. .

  The microcomputer 120 of the ECU 100 controls each part of the engine while executing various determinations and calculations based on input signals from the sensors. That is, the ECU 100 is electrically connected to the intake VVT 13a, the exhaust VVT 14a, the injector 15, the spark plug 16, the throttle valve 32, the electromagnetic clutch 34, the bypass valve 39, the EGR valve 53, and the like. Based on this, a control signal is output to each of these devices.

(3) Combustion Control (3-1) Basic Control FIG. 5 is a map for explaining the difference in control according to the engine speed / load. As shown in the figure, the engine operating region is roughly divided into four operating regions A1 to A4 depending on the combustion mode. Assuming that the first operation region A1, the second operation region A2, the third operation region A3, and the fourth operation region A4, respectively, the first operation region A1 is a low speed / low load region in which both the rotational speed and the load are low. The third operation region A3 is a low speed / high load region where the rotation speed is low and the load is high, the fourth operation region A4 is a high speed region where the rotation speed is high, and the second operation region A2 is the first operation region. The remaining regions excluding the third and fourth operation regions A1, A3, and A4 (in other words, the region combining the low speed / medium load region and the medium speed region). Hereinafter, the combustion mode selected in each operation region will be described in order.

(A) First operation region In the first operation region A1 of low speed and low load, partial compression is a mixture of SI combustion and CI combustion in a state where supercharging by the supercharger 33 is stopped (a state of natural intake). Ignition combustion (hereinafter referred to as SPCCI combustion) is performed. Note that “SPCCI” in SPCCI combustion is an abbreviation of “Spark Controlled Compression Ignition”.

  Here, SI combustion is a mode in which the air-fuel mixture is ignited by the spark plug 16 and the air-fuel mixture is forcibly burned by flame propagation that expands the combustion region from the ignition point to the surroundings. This is a form in which the air-fuel mixture is combusted by self-ignition in an environment where the temperature of the piston 5 is increased and the pressure is increased. The SPCCI combustion, which is a mixture of SI combustion and CI combustion, SI burns a part of the air-fuel mixture in the combustion chamber 6 by spark ignition performed in an environment just before the air-fuel mixture self-ignites. This is a combustion mode in which the remaining air-fuel mixture in the combustion chamber 6 is subjected to CI combustion by self-ignition after combustion (by further increase in temperature and pressure accompanying SI combustion).

  In SPCCI combustion, heat generation during SI combustion is gentler than heat generation during CI combustion. For example, in the waveform of the heat generation rate when SPCCI combustion is performed, the rising slope is relatively small as shown in FIG. Further, the pressure fluctuation in the combustion chamber 6 (that is, dP / dθ: P is the in-cylinder pressure θ is the crank angle) is also gentler during SI combustion than during CI combustion. In other words, the waveform of the heat generation rate at the time of SPCCI combustion has a first heat generation rate portion formed by SI combustion with a relatively small rising gradient and a relatively large rising gradient formed by CI combustion. The second heat generation part is formed to be continuous in this order.

  When the temperature and pressure in the combustion chamber 6 are increased by the SI combustion, the unburned mixture is self-ignited and CI combustion is started. As illustrated in FIG. 6 or FIG. 7 described later, the slope of the heat generation rate waveform changes from small to large at the timing of this self-ignition (that is, the timing at which CI combustion starts). That is, the waveform of the heat generation rate in SPCCI combustion has an inflection point (X in FIG. 7) that appears at the timing when CI combustion starts.

  After the start of CI combustion, SI combustion and CI combustion are performed in parallel. Since CI combustion generates more heat than SI combustion, the heat generation rate is relatively large. However, since CI combustion is performed after the compression top dead center, the slope of the heat generation rate waveform does not become excessive. That is, when the compression top dead center is passed, the motoring pressure is lowered due to the lowering of the piston 5, and as a result, the increase in the heat generation rate is suppressed, so that it is avoided that dp / dθ during CI combustion becomes excessive. The Thus, in SPCCI combustion, due to the nature that CI combustion is performed after SI combustion, dp / dθ, which is an indicator of combustion noise, is unlikely to be excessive, and simple CI combustion (when all fuels are subjected to CI combustion) ), Combustion noise can be suppressed.

  As the CI combustion ends, the SPCCI combustion also ends. Since CI combustion has a higher combustion speed than SI combustion, the combustion end timing can be advanced compared to simple SI combustion (when all fuels are subjected to SI combustion). In other words, in SPCCI combustion, the combustion end timing can be brought close to the compression top dead center in the expansion stroke. Thereby, in SPCCI combustion, fuel consumption performance can be improved compared with simple SI combustion.

  In order to realize the SPCCI combustion as described above, in the first operation region A1, each part of the engine is controlled by the ECU 100 as follows.

  The injector 15 injects all or most of the fuel to be injected during one cycle during the compression stroke. For example, at the operation point P1 included in the first operation region A1, the injector 15 injects fuel in two parts from the middle stage to the latter stage of the compression stroke as shown in the chart (a) of FIG.

  The spark plug 16 ignites the air-fuel mixture in the vicinity of the compression top dead center. For example, at the operating point P1, the spark plug 16 ignites the air-fuel mixture at a timing slightly ahead of the compression top dead center (chart (a) in FIG. 6).

  The supercharger 33 is turned off. That is, the electromagnetic clutch 34 is released, the connection between the supercharger 33 and the engine body 1 is released, and the bypass valve 39 is fully opened, whereby the supercharging by the supercharger 33 is stopped.

  The intake VVT 13a and the exhaust VVT 14a are set to a timing such that a sufficient valve overlap period is formed in which both the intake and exhaust valves 11 and 12 are opened across the exhaust top dead center. Thereby, internal EGR which makes burned gas remain in the combustion chamber 6 is implement | achieved, and the temperature (initial temperature before compression) of the combustion chamber 6 is raised.

  The throttle valve 32 is fully opened.

  The EGR valve 53 basically has an opening degree so that an air-fuel ratio (A / F), which is a weight ratio of air (fresh air) in the combustion chamber 6 to fuel, becomes a predetermined air-fuel ratio. Be controlled. Specifically, the air-fuel ratio in the first operating region A1 is lower than the load line L shown in FIG. 5, leaner than the theoretical air-fuel ratio (λ> 1), and higher than the load line L. Is set to the theoretical air-fuel ratio or the vicinity thereof (λ≈1). The basic opening degree of the EGR valve 53 is set to an opening degree at which this air-fuel ratio is realized. Note that λ is an excess air ratio, λ = 1 when the air-fuel ratio is the stoichiometric air-fuel ratio (14.7), and λ> 1 when the air-fuel ratio is leaner than the stoichiometric air-fuel ratio. In the first operation region A1, as described above, the air-fuel ratio (A / F) is set to the stoichiometric air-fuel ratio or leaner than that, and EGR gas (external EGR gas and internal EGR gas) is introduced into the combustion chamber 6. Therefore, the gas air-fuel ratio (G / F), which is the weight ratio of the total gas in the combustion chamber 6 to the fuel, becomes lean in any of the first operation region A1.

  The swirl valve 18 is fully closed or closed to a low opening degree close to full close so that a strong swirl flow is formed in the combustion chamber 6. This swirl flow grows during the intake stroke and remains halfway through the compression stroke. For this reason, for example, when the fuel is injected during the compression stroke as in the operation point P1 described above, the injected fuel can be collected in the central portion of the combustion chamber 6 where the swirl flow is relatively weak, and the fuel is stratified. Is realized. For example, the air-fuel ratio in the central portion of the combustion chamber 6 is set to 20 or more and 30 or less, and the air-fuel ratio in the outer peripheral portion of the combustion chamber 6 is set to 35 or more.

(B) Second operation region In the second operation region A2 (region combining the low speed / medium load region and the medium speed region), in order to cause the mixture to perform SPCCI combustion while supercharging by the supercharger 33, the following Such control is executed.

  The injector 15 injects a part of the fuel to be injected during one cycle during the intake stroke and injects the remaining fuel during the compression stroke. For example, at the operation point P2 included in the second operation region A2, the injector 15 performs the first fuel injection during the intake stroke, as shown in the chart (b) of FIG. 6, during the intake stroke. At the same time, the second fuel injection for injecting a smaller amount of fuel than the first fuel injection is performed during the compression stroke. Further, at the operation point P3 at a higher load and higher rotation side than the operation point P2, the injector 15 injects fuel over a series of periods from the intake stroke to the compression stroke as shown in the chart (c) of FIG. .

  The spark plug 16 ignites the air-fuel mixture in the vicinity of the compression top dead center. For example, the spark plug 16 ignites the air-fuel mixture at a timing slightly ahead of the compression top dead center at the operating point P2 (chart (b) in FIG. 6), and from the compression top dead center at the operating point P3. The air-fuel mixture is ignited at a slightly retarded timing (chart (c) in FIG. 6).

  The supercharger 33 is turned on. That is, supercharging by the supercharger 33 is performed by engaging the electromagnetic clutch 34 and connecting the supercharger 33 and the engine body 1. At this time, the bypass valve is set so that the pressure (supercharging pressure) in the surge tank 36 detected by the second intake pressure sensor SN6 matches the target pressure predetermined for each operating condition (rotational speed / load). The opening of 39 is controlled.

  The intake valve 11 and the exhaust VVT 13a and the exhaust VVT 14a are subjected to internal EGR only at a part of the low load side of the second operation region A2 (in other words, the internal EGR is stopped at the high load side). The valve timing of the valve 12 is controlled.

  The throttle valve 32 is fully opened.

  The opening degree of the EGR valve 53 is basically controlled so that the air-fuel ratio (A / F) in the combustion chamber 6 becomes a predetermined air-fuel ratio. Specifically, the air-fuel ratio in the second operation region A2 is leaner (λ> 1) on the lower load side than the load line L (FIG. 5), or the stoichiometric air-fuel ratio or higher on the higher load side than the load line L. It is set in the vicinity (λ≈1). For example, the recirculation amount of the exhaust gas is adjusted so as to decrease as the load increases, and is almost zero in the vicinity of the maximum load of the engine. In other words, the gas air-fuel ratio in the combustion chamber 6 is all lean except in the vicinity of the maximum engine load.

  The swirl valve 18 is fully closed, or is opened to an appropriate intermediate opening degree other than full close / full open. Specifically, the swirl valve 18 is fully closed in a part on the low load side of the second operation region A2, and is set to an intermediate opening degree in the remaining region on the high load side. In addition, the opening degree of the swirl valve 18 in the latter area | region is enlarged, so that load is high.

(C) Third operation region In the third operation region A3 of low speed and high load, the following control is executed to perform SI combustion of the air-fuel mixture while supercharging by the supercharger 33.

  The injector 15 injects a part of the fuel to be injected during one cycle during the intake stroke and injects the remaining fuel during the compression stroke. For example, at the operation point P4 included in the third operation region A3, the injector 15 performs the first fuel injection during the intake stroke, as shown in the chart (d) of FIG. 6, during the intake stroke. At the same time, the second fuel injection for injecting a smaller amount of fuel than the first fuel injection is executed in the latter half of the compression stroke (just before the compression top dead center).

  The spark plug 16 ignites the air-fuel mixture at a relatively late timing, for example, when about 5 to 20 ° CA has elapsed from the compression top dead center. Then, the SI combustion is started by this ignition, and all of the air-fuel mixture in the combustion chamber 6 is combusted by flame propagation.

  The supercharger 33 is turned on. At this time, the opening degree of the bypass valve 39 is controlled so that the pressure in the surge tank 36 (supercharging pressure) matches the target pressure.

  The throttle valve 32 is fully opened.

  The opening degree of the EGR valve 53 is controlled so that the air-fuel ratio (A / F) in the combustion chamber 6 becomes the stoichiometric air-fuel ratio or slightly richer. On the other hand, the gas air-fuel ratio (G / F) in the combustion chamber 6 is lean except for the vicinity of the maximum engine load.

  The opening degree of the swirl valve 18 is set to a predetermined intermediate opening degree (for example, 50%) or a value close thereto.

(D) Fourth operation region In the fourth operation region A4 on the higher speed side than the first to third operation regions A1 to A3, the following control is performed so that relatively orthodox SI combustion is realized. Is called.

  The injector 15 injects at least a predetermined period that overlaps with the intake stroke. For example, at the operation point P5 included in the fourth operation region A4, the injector 15 injects fuel over a series of periods from the intake stroke to the compression stroke, as shown in the chart (e) of FIG.

  The spark plug 16 ignites the air-fuel mixture in the vicinity of the compression top dead center. For example, at the operation point P5, the spark plug 16 ignites the air-fuel mixture at a timing slightly ahead of the compression top dead center (chart (e) in FIG. 8). Then, the SI combustion is started by this ignition, and all of the air-fuel mixture in the combustion chamber 6 is combusted by flame propagation.

  The supercharger 33 is turned on. At this time, the opening degree of the bypass valve 39 is controlled so that the pressure in the surge tank 36 (supercharging pressure) matches the target pressure.

  The throttle valve 32 is fully opened.

  The opening degree of the EGR valve 53 is controlled so that the air-fuel ratio (A / F) in the combustion chamber 6 becomes the stoichiometric air-fuel ratio or slightly richer. On the other hand, the gas air-fuel ratio (G / F) in the combustion chamber 6 is lean except for the vicinity of the maximum engine load.

  The swirl valve 18 is fully opened. Thereby, not only the first intake port 9A but also the second intake port 9B is completely opened, and the charging efficiency of the engine is increased.

(3-2) SI Rate As described above, in this embodiment, SPCCI combustion in which SI combustion and CI combustion are mixed is executed in the first operation region A1 and the second operation region A2, but in this SPCCI combustion, It is important to control the ratio of SI combustion and CI combustion according to the operating conditions.

  Here, in this embodiment, the SI rate that is the ratio of the heat generation amount by SI combustion to the total heat generation amount by SPCCI combustion (SI combustion and CI combustion) is used as the ratio. FIG. 7 is a diagram for explaining the SI rate, and shows a change in the heat generation rate (J / deg) depending on the crank angle when the SPCCI combustion occurs. The inflection point X in the waveform of FIG. 7 is an inflection point that appears when the combustion mode is switched from SI combustion to CI combustion, and the crank angle θci corresponding to this inflection point X is defined as the start timing of CI combustion. can do. Then, the area Q1 of the waveform of the heat generation rate located on the advance side with respect to θci (CI combustion start timing) is defined as the heat generation amount by SI combustion, and the waveform of the heat generation rate located on the retard side with respect to θci. The area Q2 is defined as the heat generation rate for CI combustion. Thus, the above-described SI rate defined by (heat generation amount by SI combustion) / (heat generation amount by SPCCI combustion) is expressed as SI rate = Q1 / (Q1 + Q2) using the areas Q1 and Q2. be able to.

  In the case of CI combustion, the air-fuel mixture is simultaneously and frequently burned by self-ignition, so that the heat generation rate is likely to be higher than that of SI combustion by flame propagation, and large noise is likely to be generated. For this reason, it is desirable that the SI rate (= Q1 / (Q1 + Q2)) in SPCCI combustion generally increases as the load increases. This is because when the load is high, the amount of fuel injected is large and the total amount of heat generated in the combustion chamber 6 is large, so the SI rate is reduced (that is, the rate of CI combustion is increased). This is because a large noise is generated. On the contrary, since CI combustion is excellent in terms of thermal efficiency, it is preferable to perform CI combustion with as much fuel as possible unless noise becomes a problem. For this reason, it is desirable that the SI rate in SPCCI combustion is generally smaller as the load is lower (that is, the rate of CI combustion is increased). From this point of view, in this embodiment, the target SI rate (target SI rate) is determined in advance according to the engine operating conditions so that the SI rate increases as the load increases. Control value target values such as ignition timing, fuel injection amount / injection timing, and in-cylinder state amount are determined so that the rate is realized. Note that the in-cylinder state quantity referred to here is, for example, the temperature in the combustion chamber 6 or the EGR rate. The EGR rate includes the external EGR rate that is the ratio of the external EGR gas (exhaust gas recirculated to the combustion chamber 6 through the EGR passage 51) to the total gas in the combustion chamber 6, and the internal EGR to the total gas in the combustion chamber 6. The internal EGR rate which is the ratio of gas (burned gas remaining in the combustion chamber 6) is included.

  For example, as the ignition timing is advanced, more fuel is combusted by SI combustion, and the SI rate becomes higher. Further, as the fuel injection timing is advanced, more fuel is burned by CI combustion, and the SI rate is lowered. Alternatively, the higher the temperature of the combustion chamber 6, the more fuel is burned by CI combustion, and the SI rate is lowered.

  Based on the above-mentioned tendency, in this embodiment, the target SI rate described above can be realized by the target values of the ignition timing, the fuel injection amount / injection timing, and the in-cylinder state amount (temperature, EGR rate, etc.). It is determined in advance for each operating condition so as to be a simple combination. During operation by SPCCI combustion (that is, during operation in the first and second operation regions A1 and A2), the ECU 100 determines the injector 15, the spark plug 16, the EGR valve 53, the intake / exhaust VVT 13a, according to the target values of these control amounts. 14a and the like are controlled. For example, the ignition plug 16 is controlled according to the target value of the ignition timing, and the injector 15 is controlled according to the target value of the fuel injection amount / injection timing. Further, the EGR valve 53 and the intake / exhaust VVTs 13a, 14a are controlled in accordance with the target values of the temperature of the combustion chamber 6 and the EGR rate, and the recirculation amount of exhaust gas (external EGR gas) through the EGR passage 51 and burned by internal EGR. The residual amount of gas (internal EGR gas) is adjusted. That is, the spark plug, the injector 15, the EGR valve 53, and the intake / exhaust VVTs 13a and 14a are controlled according to the basic control described in (3-1), and further controlled so that the target SI rate is realized in detail. Is done.

(4) Signal Processing of In-Cylinder Pressure Sensor (4-1) Pressure Conversion Processing FIG. 8 is a diagram for explaining signal processing of the in-cylinder pressure sensor. As shown in FIG. 8, the I / F circuit 110 of the ECU 100 is provided with a low-pass filter (LPF) 111. Further, the microcomputer 120 of the ECU 100 is provided with an A / D converter 121, and functionally includes a pressure conversion unit (correction unit) 122, a combustion gravity center timing calculation unit (combustion gravity center timing estimation unit) 123, and the like. Is provided.

  The signal from the in-cylinder pressure sensor SN2 is first input to the low-pass filter 111 of the I / F circuit 110. The low-pass filter 111 is a filter that outputs only a waveform having a predetermined frequency or less. By passing through the low-pass filter 111, high-frequency electrical noise (so-called white noise) is removed from the signal of the in-cylinder pressure sensor SN2. The signal of the in-cylinder pressure sensor SN2 output from the low-pass filter 111 is input to the A / D converter 121 and converted into a digital signal by the A / D converter 121. The signal of the in-cylinder pressure sensor SN2 converted into the digital signal is input to the pressure conversion unit 122.

  The pressure converter 122 is a part that converts the signal of the in-cylinder pressure sensor SN2 into pressure. That is, as described above, the signal output from the in-cylinder pressure sensor SN2 is a voltage signal corresponding to the in-cylinder pressure, and the voltage conversion unit 122 converts this voltage signal into pressure.

  Here, the relationship between the output value (voltage value) of the in-cylinder pressure sensor SN2 and the pressure can be expressed by pressure = output value (voltage value) of the in-cylinder pressure sensor SN2 × proportional coefficient + offset amount. And this proportionality coefficient is decided beforehand. On the other hand, the offset amount is determined to some extent, but the exact value differs for each in-cylinder pressure sensor SN2. Therefore, in the present embodiment, this offset amount is calculated using the value of the intake pressure sensor SN6, and the output value (voltage value) of the in-cylinder pressure sensor SN2 is converted into pressure. That is, the value obtained by multiplying the output value (voltage value) of the in-cylinder pressure sensor SN2 by the proportional coefficient is corrected to the actual pressure using the value of the intake pressure sensor SN6.

  In the present embodiment, the conversion process from voltage to pressure (hereinafter simply referred to as conversion process) is performed individually for each in-cylinder pressure sensor SN2. Moreover, this conversion process is performed for every combustion cycle. Specifically, the output value of each in-cylinder pressure sensor SN2 (output value after digital processing) is individually stored in a storage unit (not shown) of the ECU 100 over a predetermined period. The conversion process is performed on the output value for one combustion cycle of each in-cylinder pressure sensor SN2.

  The procedure of the conversion process will be described with reference to the flowchart of FIG.

  First, in step S1, the pressure conversion unit 122 reads an intake valve closing timing IVC that is a timing at which the intake valve 11 is closed. Specifically, the intake valve closing timing IVC of the cylinder 2 provided with the in-cylinder pressure sensor SN2 to be converted and the intake valve closing timing IVC in the combustion cycle to be converted are read.

  Next, in step S2, the pressure conversion unit 122 determines whether or not the intake valve closing timing IVC is a timing retarded from the intake bottom dead center.

  If the determination in step S2 is NO and the intake valve closing timing IVC is the intake bottom dead center or a timing that is further advanced than this, the pressure conversion unit 122 proceeds to step S3. In step S3, the pressure conversion unit 122 sets a period from a timing before a predetermined crank angle to the intake valve closing timing IVC before the intake valve closing timing IVC as an average processing period (reference period).

  On the other hand, when the determination in step S2 is YES and the intake valve closing timing IVC is a timing retarded from the intake bottom dead center, the pressure conversion unit 122 proceeds to step S4. In step S4, the pressure conversion unit 122 sets a period from the timing before a predetermined crank angle before the intake bottom dead center to the intake bottom dead center as an average processing period (reference period).

  After step S3 or step S4, the process proceeds to step S5.

  In step S5, the pressure conversion unit 122 reads a plurality of intake pressures detected by the intake pressure sensor SN6 during the average processing period set in step S3 or step S4, and in-cylinder pressure sensor SN2 during the average processing period. A plurality of voltage values output from are read from the storage unit. The storage unit stores an output value from the intake pressure sensor SN6 over a predetermined period including at least the average processing period.

  The average processing period is set to 12 ° CA (crank angle), for example, and the output of the in-cylinder pressure sensor SN2 is stored in the storage unit in units of 3 ° CA, for example. Accordingly, in step S2, for example, outputs (voltage values) of four in-cylinder pressure sensors SN2 are read.

  Next, in step S6, the pressure conversion unit 122 calculates the average value Pim_ave of the plurality of intake pressures read in step S5, that is, the average value Pim_ave of the intake pressure in the average processing period.

  In step S7, the pressure conversion unit 122 determines the average value Vcps_ave of the output values (voltage values) of the plurality of in-cylinder pressure sensors SN2 read in step S5, that is, the output value (voltage value) of the in-cylinder pressure sensor SN2 in the average processing period. ) Average value Vcps_ave.

  Next, in step S8, the pressure converter 122 calculates a value obtained by multiplying the average value Vcps_ave of the output value (voltage value) of the in-cylinder pressure sensor SN2 calculated in step S7 by the proportional coefficient K before offset correction. Calculated as internal pressure. That is, in step S8, the cylinder pressure before offset correction is set to Pcps_of, and this is calculated by Pcps_of = K × Vcps_ave.

  Next, in step S9, the pressure conversion unit 122 calculates, as an offset amount, a value obtained by subtracting the pre-offset cylinder pressure Pcps_of calculated in step S8 from the average value Pim_ave of the intake pressure calculated in step S6. That is, the offset amount is set as OFFSET, and this is calculated by OFFSET = Pim_ave−Pcps_of.

  Next, in step S10, the pressure conversion unit 122 converts the output value (voltage value) of the in-cylinder pressure sensor SN2 stored in the storage unit into a pressure using the proportional coefficient K and the offset amount OFFSET, respectively. . Specifically, the in-cylinder pressure is Pcps, and this is calculated by Pcps = K × (output value (voltage value) of in-cylinder pressure sensor SN2) + OFFSET.

  Thus, in this embodiment, first, the detected value of the in-cylinder pressure sensor SN2 output as a voltage is converted into the in-cylinder pressure before offset correction as shown by the broken line in FIG. Then, the in-cylinder pressure before offset correction is corrected so that the average value of the in-cylinder pressure and the average value of the intake pressure in the average processing period coincide with each other, and thereby the output value (voltage value) of the in-cylinder pressure sensor SN2 is a solid line in FIG. Is converted into in-cylinder pressure (pressure) as shown in FIG.

(4-2) Combustion parameter calculation (a) Heat generation amount The method of using the in-cylinder pressure detected by the in-cylinder pressure sensor obtained as described above is not particularly limited. In 123, it is calculated using the in-cylinder pressure at which the combustion center-of-gravity timing, which is a parameter related to the amount of heat generation and combustion, is detected. In the combustion center-of-gravity period, the amount of heat generated (the amount of heat generated in the combustion chamber 6 by combustion during one combustion cycle) reaches 50% of the total amount of heat generated in the combustion chamber 6 during one combustion cycle. It's time.

  Specifically, the combustion center-of-gravity timing calculation unit 123 first calculates the heat generation rate (heat generation speed) ΔQ (θ) at the crank angle θ and the in-cylinder pressure P (pressure) detected by the in-cylinder pressure sensor SN2. And calculated from the following equation (1). In the present embodiment, as described above, the output of the in-cylinder pressure sensor SN2 is stored in 3 ° CA increments in the storage unit of the ECU 100, and here, the heat generation rate per 3 ° CA is calculated in 3 ° CA increments. Is done.

  In Equation (1), κ is the specific heat ratio of the gas in the combustion chamber 6 and a preset value is used. V (θ) is the volume of the combustion chamber 6 at the crank angle θ. The ECU 100 stores the volume of the combustion chamber 6 with respect to each crank angle as a map, and the volume V (θ) of the combustion chamber 6 corresponding to the crank angle θ is extracted from this map and used in the equation (1). ΔV (θ) is the rate of change of the volume V of the combustion chamber 6 at the crank angle θ. Here, the combustion chamber 3 ° CA before the crank angle θ from the volume V of the combustion chamber 6 at the crank angle θ. A value obtained by subtracting the volume V of 6 is used. That is, a value calculated by ΔV (θ) = V (θ) −V (θ−3 ° CA) is used for ΔV (θ).

  P (θ) is the in-cylinder pressure P at the crank angle θ (the in-cylinder pressure (pressure) at the crank angle θ obtained from the in-cylinder pressure sensor SN2). ΔP (θ) is a changing speed of the in-cylinder pressure P at the crank angle θ, and here, a value obtained by subtracting the in-cylinder pressure P 3 ° CA before the crank angle θ from the in-cylinder pressure P at the crank angle θ. Used. That is, a value calculated by ΔP (θ) = P (θ) −P (θ−3 ° CA) is used for ΔP (θ).

  Next, the combustion center-of-gravity timing calculation unit 123 integrates the calculated heat generation rate to calculate the heat generation amount Q (θ) at each crank angle θ. Specifically, the heat generation amount Q (θ) is calculated by the following equation (2).

  In the present embodiment, Δθ is set to 3 ° CA as described above. Further, in the present embodiment, only the heat generation rate and the heat generation amount in a predetermined period set to include a period in which combustion occurs are calculated. Then, the heat generation amount Q (θ_s) at the start timing θ_s of this period, that is, the initial value Q (θ_s) of the heat generation amount serving as a reference for the integrated value is the in-cylinder pressure P (in-cylinder pressure sensor SN2) at the start timing θ_s. Is calculated from the difference between the reference pressure and the in-cylinder pressure (pressure) at the start time θ_s obtained from the above. The reference pressure is the in-cylinder pressure at the start timing θ_s when combustion does not occur, that is, the in-cylinder pressure at the start timing θ_s during motoring. This reference pressure is obtained by experiments or the like and stored in the ECU 100. Specifically, the reference pressure is stored in the ECU 100 as a map or the like for each of a plurality of engine speeds and a plurality of engine loads, and the ECU 100 generates a current (a heat generation amount Q (θ) to be calculated). The reference pressure corresponding to the engine speed and the engine load is extracted and used for the calculation of equation (2).

  Next, the combustion center-of-gravity timing calculation unit 123 extracts the maximum value and the minimum value from the calculated heat generation amount Q (θ), and the average value thereof, that is, a value that is half the sum of the maximum value and the minimum value. Ask for. Then, the crank angle θ that is closest to the average value is extracted from the heat generation amount Q (θ) at each crank angle θ, and the crank angle θ is determined as the combustion center-of-gravity timing.

  The combustion center-of-gravity timing calculated in this way is used for ignition timing control and the like. Specifically, as described above, in this embodiment, each unit is controlled according to the basic control described in (3-1), and further, the target SI rate described in (3-2) is realized. Basically, each part is controlled so that SPCCI combustion or SI combustion is appropriately realized in each region, and the target SI rate is realized in the region where SPCCI combustion is performed. In the embodiment, in addition to this, the ignition timing is corrected so that the combustion center-of-gravity timing calculated by the combustion center-of-gravity timing calculation unit 123 as described above falls within a predetermined range.

(5) Operation, etc. As described above, in the present embodiment, the output voltage of the in-cylinder pressure sensor SN2 is based on the intake pressure detected by the intake pressure sensor SN6 during the average processing period set near the intake bottom dead center. The final in-cylinder pressure is calculated by correcting the in-cylinder pressure calculated by applying the proportional coefficient K, that is, the in-cylinder pressure before offset correction. Therefore, the in-cylinder pressure detected by the in-cylinder pressure sensor SN2 can be easily corrected to an in-cylinder pressure closer to the actual value.

  Specifically, in the vicinity of the intake bottom dead center, the pressure in the cylinder 2 and the pressure in the intake passage 30 are substantially the same. Therefore, as described above, if the detection value of the in-cylinder pressure sensor SN2 is corrected using the pressure in the intake passage 30 near the bottom dead center of the intake air detected by the intake pressure sensor SN6, the actual pressure is more reliably approached. be able to.

  In particular, in the present embodiment, in the engine in which the intake passage 30 communicates with the plurality of cylinders 2, the in-cylinder pressure is calculated using the average value of the plurality of intake pressures detected by the intake pressure sensor SN6 during the average processing period. It is corrected. Therefore, it is possible to prevent the correction from being appropriately performed due to the pulsation of the intake pressure. Specifically, when the intake passage 30 communicates with the plurality of cylinders 2, the pressure in the intake passage 30, that is, the intake pressure pulsates, even in the vicinity of the bottom dead center of the intake, depending on the timing, There is a risk that the pressure and the intake pressure may deviate. On the other hand, when the average value of the intake pressure is used for correction as described above, the in-cylinder pressure can be corrected more appropriately while suppressing this deviation.

  Further, in the present embodiment, when the intake valve closing timing IVC is a timing retarded from the intake bottom dead center, a period from a timing before a predetermined crank angle to the intake bottom dead center before the intake bottom dead center. Is set as the average processing period, and the in-cylinder pressure is corrected based on the intake pressure detected by the intake pressure sensor SN6 until the intake bottom dead center. Therefore, the in-cylinder pressure can be corrected more appropriately.

  Specifically, if the intake valve 11 is still open after the intake bottom dead center, the intake air blows back from the cylinder 2 into the intake passage 30, and a difference occurs between the pressure in the intake passage 30 and the pressure in the cylinder 2. There is a case. Specifically, the pressure in the intake passage 30 may be higher than the pressure in the cylinder 2. For this reason, if the in-cylinder pressure is corrected using the intake pressure in a state where there is the deviation, the in-cylinder pressure may also deviate from an actual value. On the other hand, in the present embodiment, as described above, when the intake valve closing timing IVC is a timing retarded from the intake bottom dead center and the intake air blows back, the intake valve bottom timing is lower than the intake bottom dead center. The period from the time before a predetermined crank angle to the intake bottom dead center and the period during which intake air does not blow back is set as the average processing period, and the in-cylinder pressure is corrected using the intake pressure during this period. Accordingly, the in-cylinder pressure can be accurately corrected while suppressing the deviation between the in-cylinder pressure and the intake pressure during the average processing period for performing the correction.

  Further, when the intake valve closing timing IVC is the intake bottom dead center or a timing that is more advanced than the intake bottom dead center, the intake valve closing timing IVC is from a timing before a predetermined crank angle to the intake valve closing timing IVC. This period is set as the average processing period. Therefore, the in-cylinder pressure can be corrected more appropriately. That is, after the intake valve 11 is closed, the pressure in the cylinder 2 and the pressure in the intake passage 30 may be greatly deviated. On the other hand, in this embodiment, when the intake valve closing timing IVC is an intake bottom dead center or a timing that is on the more advanced side than the intake bottom dead center, the intake valve closing timing IVC is a predetermined crank angle before the intake valve closing timing IVC. The period from the timing to the intake valve closing timing IVC and before the intake valve 11 is closed is set as an average processing period, and the in-cylinder pressure is corrected using the intake pressure during this period. Therefore, the in-cylinder pressure can be accurately corrected while suppressing the deviation.

  According to the present embodiment, as described above, the in-cylinder pressure can be accurately corrected to obtain a value close to the actual in-cylinder pressure, so that the combustion center-of-gravity timing can be calculated more accurately. The combustion state of the air-fuel mixture in the combustion chamber 6 can be made appropriate by controlling the combustion center of gravity timing to an appropriate time. In particular, when used in an engine that performs SPCCI combustion as in this embodiment, the controllability of adjustment of the combustion center of gravity is improved, and the SI rate can be adjusted more appropriately according to the operating state of the engine. Therefore, the fuel efficiency effect and combustion noise suppression effect in SPCCI combustion can be further enhanced.

(6) Modified Example In the above embodiment, the case where the ignition timing is corrected so that the combustion center-of-gravity timing falls within a predetermined range has been described, but this control may be omitted.

  Further, the method of using the combustion center-of-gravity time is not limited to the above. For example, when the combustion noise is estimated using the signal of the in-cylinder pressure sensor SN2, the period for performing the estimation may be set with the combustion gravity center time as a reference.

  Moreover, although the said embodiment demonstrated the case where the in-cylinder pressure (pressure) obtained by converting the detected value of in-cylinder pressure sensor SN2 was used for calculating | requiring a combustion gravity center time and a combustion noise parameter | index, in-cylinder pressure The method of using (pressure) is not limited to this.

  For example, in an engine in which the ignition timing of SI combustion or CI combustion is calculated using the heat generation rate and the ignition timing is corrected using this ignition timing, the in-cylinder pressure is used for calculating the heat generation rate and the ignition timing. Absolute pressure may be used. Further, in an engine configured to keep the in-cylinder pressure below a predetermined value, the absolute value of the in-cylinder pressure obtained from the in-cylinder pressure sensor SN2 is used, and the fuel injection amount or The ignition timing and the like may be controlled.

  The number of cylinders in the engine body is not limited to the above.

DESCRIPTION OF SYMBOLS 1 Engine main body 2 Cylinder 6 Combustion chamber 11 Intake valve 122 Pressure conversion part (correction means)
123 Combustion center of gravity timing calculation unit (combustion center of gravity timing estimation means)
SN2 In-cylinder pressure sensor (in-cylinder pressure detection means)
SN6 Intake pressure sensor (Intake pressure detection means)

Claims (5)

A signal processing device provided in an engine comprising an engine body in which a cylinder is formed, and an intake passage through which intake air introduced into the cylinder flows.
In-cylinder pressure detecting means for detecting an in-cylinder pressure that is a pressure in the cylinder;
An intake pressure detecting means for detecting an intake pressure which is a pressure in the intake passage;
Correction means for correcting the in-cylinder pressure detected by the in-cylinder pressure detection means based on the intake pressure detected by the intake pressure detection means during a reference period set near the intake bottom dead center. An engine signal processing device. The engine signal processing device according to claim 1,
The engine body has a plurality of the cylinders,
The intake passage communicates with the cylinders,
The correction means calculates an average value of a plurality of intake pressures detected by the intake pressure detection means during the reference period, and corrects the in-cylinder pressure detected by the in-cylinder pressure detection means using the average value. A signal processing apparatus for an engine. The signal processing apparatus for an engine according to claim 1 or 2,
The engine signal processing apparatus according to claim 1, wherein the reference period is set to a period earlier than a closing timing of an intake valve that opens and closes an intake port communicating with the cylinder. The engine signal processing device according to claim 3,
The correction means sets the reference period to a period earlier than the intake bottom dead center when the closing timing of the intake valve is retarded from the intake bottom dead center. Signal processing equipment. An engine signal processing device according to any one of claims 1 to 4,
Based on the in-cylinder pressure after being corrected by the correcting means, the amount of heat generated in the cylinder is estimated, and the center of combustion is the point in time when 50% of the fuel supplied to the cylinder burns in one combustion cycle. An engine signal processing device comprising combustion centroid timing estimation means for estimating timing.