JP2017020414A - Control device for internal combustion engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a control device for an internal combustion engine that can accurately estimate an EGR rate while successfully reflecting temperature characteristics of a specific heat ratio on the basis of a state change of an air-fuel mixture in a compression stroke and thus can appropriately control the internal combustion engine.SOLUTION: In this control device for an internal combustion engine, a reference crank angle CA_REF immediately before combustion of an air-fuel mixture is started is set in accordance with an operating state of the internal combustion engine 3 (Step 5), and in the condition of the air-fuel mixture where the EGR rate is equal to a target EGR rate and an air-fuel ratio is a theoretical air-fuel ratio, reference cylinder inner pressure P_REF generated at the reference crank angle CA_REF is calculated on the basis of temperature characteristics of a specific heat ratio of the air-fuel mixture (Step 6). On the basis of a pressure difference ΔP between an actual cylinder inner pressure P_CPS detected by a cylinder inner pressure sensor 51 at the reference crank angle CA_REF and a reference cylinder inner pressure P_REF, an EGR rate R_EGR is estimated (Steps 9-11). In accordance with the EGR rate R_EGR, the internal combustion engine 3 is controlled (Figure 18).SELECTED DRAWING: Figure 3

Description

  The present invention relates to a control device for an internal combustion engine including an EGR device that recirculates a part of exhaust gas discharged from the internal combustion engine to an intake passage, and more particularly to a control device that estimates an EGR rate and controls the internal combustion engine according to the EGR rate. About.

  As a conventional method for estimating the EGR rate, the amount of fresh air sucked into the cylinder is detected by an air flow sensor, and the total amount of gas sucked into the cylinder is calculated based on the intake pressure detected by the intake pressure sensor. It is known to estimate the EGR rate from the fresh air amount and the total gas amount. However, when this estimation method is used in a low-pressure EGR device (an EGR device that extracts exhaust gas from the downstream side of the turbine of the turbocharger and returns it to the upstream side of the compressor of the intake passage), the external EGR gas up to the cylinders Since the flow path is relatively long and the delay until the external EGR gas reaches the cylinder is large, it is difficult to accurately estimate the EGR rate.

Further, as another conventional method for estimating the EGR rate, for example, one disclosed in Patent Document 1 is known. This estimation method focuses on the fact that the change in the state of the air-fuel mixture during the compression stroke of the internal combustion engine is a polytropic change, and that the specific heat ratio changes according to the composition of the air-fuel mixture. Specifically, in-cylinder pressures P1 and P2 are detected by in-cylinder pressure sensors at two predetermined crank angles CA1 and CA2 during the compression stroke, and cylinder volumes corresponding to these in-cylinder pressures P1 and P2 and crank angles CA1 and CA2 are detected. From V1 and V2, the specific heat ratio κ of the air-fuel mixture is calculated by the following equation.
κ = log (P1 / P2) / log (V2 / V1)
Then, the EGR rate is calculated by referring to a predetermined table that defines the relationship between the specific heat ratio and the EGR rate (EGR gas concentration) based on the calculated specific heat ratio κ.

JP 2008-231995 A

  In the above-described conventional EGR rate estimation method, it is assumed that the state change of the air-fuel mixture in the compression stroke is a polytropic change, and the air-fuel mixture is determined based on the amount of change in the in-cylinder pressures P1 and P2 detected during the compression stroke. The specific heat ratio is calculated. On the other hand, in actual control of the internal combustion engine, the ignition timing is usually changed according to the load and the rotational speed of the internal combustion engine, and the start timing of combustion of the air-fuel mixture is usually changed accordingly.

  On the other hand, in the conventional estimation method, since the predetermined crank angles CA1 and CA2 are uniformly set as the detection timings of the in-cylinder pressures P1 and P2, the rear crank angle CA2 is used to actually start the combustion of the air-fuel mixture. It may be later than the timing. In that case, since the change in the state of the mixture does not follow the change in the polytropy due to the pressure increase due to the combustion of the mixture, the accuracy of calculating the specific heat ratio is reduced. In order to avoid such a problem, for example, when the rear crank angle CA2 is set to a more advanced side, there is a possibility that a sufficient pressure difference between the in-cylinder pressures P1 and P2 cannot be secured. The specific heat ratio cannot be calculated with high accuracy.

  Further, since the specific heat ratio of the air-fuel mixture has a temperature characteristic that changes according to the temperature, even if the composition and EGR rate of the air-fuel mixture are the same, they change as the temperature of the air-fuel mixture increases due to compression in the compression stroke. On the other hand, in the conventional estimation method, neither the specific heat ratio nor the temperature characteristic of the specific heat ratio is taken into consideration when calculating the specific heat ratio or the EGR rate from the specific heat ratio. The EGR rate cannot be calculated accurately.

  The present invention has been made to solve the above-described problems, and can accurately estimate the EGR rate while favorably reflecting the temperature characteristics of the specific heat ratio based on the change in the state of the air-fuel mixture in the compression stroke. Another object of the present invention is to provide a control device for an internal combustion engine that can appropriately control the internal combustion engine using the estimated EGR rate.

  In order to achieve the above object, the invention according to claim 1 is directed to injecting fuel directly into the cylinder 3a and using a part of the exhaust gas discharged from the cylinder 3a to the exhaust passage 7 as an external EGR gas. 6 is a control device for an internal combustion engine that includes an EGR device 14 that recirculates to 6 and is a target EGR rate setting means (ECU2, ECU2) that sets a target EGR rate TGT_EGR that is a target of the EGR rate R_EGR of the air-fuel mixture charged in the cylinder 3a. 20), an in-cylinder pressure sensor 51 for detecting the pressure in the cylinder 3a as an in-cylinder pressure PCYL, and an operating state detecting means for detecting the operating state (ignition timing IGLOG, intake pressure PBA, engine speed NE) of the internal combustion engine 3. (The intake pressure sensor 56, the crank angle sensor 52, the ECU 2) and the air-fuel mixture filled in the cylinder 3a according to the detected operating state of the internal combustion engine 3. Reference crank angle setting means (ECU 2, step 5 in FIG. 3, FIG. 4) for setting the crank angle immediately before the start of firing as the reference crank angle CA_REF, and the target EGR rate in which the EGR rate R_EGR of the air-fuel mixture is set Based on the temperature characteristic of the specific heat ratio of the air-fuel mixture under the condition that the air-fuel ratio of the air-fuel mixture is equal to TGT_EGR and the air-fuel ratio of the air-fuel mixture is the stoichiometric air-fuel ratio, the pressure in the cylinder 3a generated at the set reference crank angle CA_REF is Reference in-cylinder pressure calculation means (ECU 2, step 6 in FIG. 3, FIG. 6) that calculates as P_REF, and actual in-cylinder pressure P_CPS detected by the in-cylinder pressure sensor 51 at the reference crank angle CA_REF and the calculated reference in-cylinder pressure P_REF Based on the pressure difference ΔP, EGR rate estimation means (ECU2, FIG. 3) for estimating the EGR rate R_EGR of the air-fuel mixture Steps 9 to 11) and control means (ECU 2, FIG. 18) for controlling the internal combustion engine 3 in accordance with the estimated EGR rate R_EGR.

  In this internal combustion engine, fuel is directly injected into the cylinder, and part of the exhaust gas discharged from the cylinder to the exhaust passage returns to the intake passage as external EGR gas. In the control device for an internal combustion engine of the present invention, a target EGR rate that is a target of the EGR rate of the air-fuel mixture is set, and the in-cylinder pressure (pressure in the cylinder) is detected by the in-cylinder pressure sensor. Further, according to the detected operating state of the internal combustion engine, the crank angle immediately before the start of combustion of the air-fuel mixture is set as the reference crank angle, and the pressure in the cylinder generated at this reference crank angle is Calculated as the reference in-cylinder pressure. The calculation of the reference in-cylinder pressure is performed based on the temperature characteristic of the specific heat ratio of the air-fuel mixture under the condition of the air-fuel mixture composition in which the EGR rate is equal to the target EGR rate and the air-fuel ratio is the stoichiometric air-fuel ratio.

  As described above, the specific heat ratio of the air-fuel mixture basically has a temperature characteristic that is determined according to the composition of the air-fuel mixture and changes according to the temperature of the air-fuel mixture. Therefore, by calculating the reference in-cylinder pressure based on the temperature characteristics of the specific heat ratio under the condition that the EGR rate of the air-fuel mixture is equal to the target EGR rate and the air-fuel ratio of the air-fuel mixture is the stoichiometric air-fuel ratio, The reference in-cylinder pressure can be determined uniquely and appropriately while reflecting the temperature characteristics of the specific heat ratio.

  According to the present invention, the in-cylinder pressure detected at the reference crank angle is obtained as the actual in-cylinder pressure, and the EGR rate is estimated based on the pressure difference between the actual in-cylinder pressure and the reference in-cylinder pressure. The actual in-cylinder pressure reflects the actual composition of the air-fuel mixture containing the external EGR gas and the temperature and specific heat ratio corresponding to the actual composition. Therefore, the EGR rate can be estimated based on the pressure difference between the actual in-cylinder pressure and the reference in-cylinder pressure.

  The reference crank angle is a crank angle immediately before the start of combustion of the air-fuel mixture, and is set according to the detected operating state of the internal combustion engine. With this reference crank angle setting, the actual in-cylinder pressure is acquired while combustion is not yet performed and the change in the state of the air-fuel mixture is maintained in the polytropic change, and a large pressure between the actual in-cylinder pressure and the reference in-cylinder pressure is obtained. A difference can be secured. Therefore, based on this pressure difference, the EGR rate can be accurately estimated while favorably reflecting the temperature characteristic of the specific heat ratio. In addition, since the internal combustion engine is controlled according to the EGR rate estimated with high accuracy, the internal combustion engine can be controlled appropriately.

  The invention according to claim 2 is the control device for the internal combustion engine according to claim 1, wherein the EGR device 14 is provided in the intake passage 6 from the downstream side of the turbine 23 of the supercharger (turbocharger 13) in the exhaust passage 7. An external EGR gas is recirculated upstream of the compressor 21 of the supercharger.

  The EGR device configured as described above is a so-called low pressure EGR device, and since the flow path of the external EGR gas is relatively long, the delay until the external EGR gas reaches the cylinder tends to be large. As described above, in the present invention, the EGR rate is estimated using the actual in-cylinder pressure, which is the actual pressure in the cylinder, and the reference in-cylinder pressure as parameters. Therefore, even in the case of a low-pressure EGR device, the delay of the external EGR gas Therefore, the EGR rate can be estimated with high accuracy without being affected by the above-described effects, and therefore, the effect of the present invention can be obtained particularly effectively.

  According to a third aspect of the present invention, in the internal combustion engine control apparatus according to the first or second aspect, the operating state detecting means includes the ignition timing IGLOG and the pressure of the intake air sucked into the cylinder 3a as the operating state of the internal combustion engine 3. (Intake pressure PBA) and the rotational speed NE of the internal combustion engine 3 are detected, and the reference crank angle setting means determines the reference crank angle CA_REF according to the detected ignition timing IGLOG, intake air pressure and the rotational speed NE of the internal combustion engine 3. Is set (FIG. 4).

  As described above, the reference crank angle is set as the crank angle immediately before the start of combustion of the air-fuel mixture. In addition, the combustion start timing of the air-fuel mixture is directly affected by the ignition timing and changes according to the intake pressure. When expressed by the crank angle, it changes according to the rotational speed of the internal combustion engine. To do. According to this configuration, since the reference crank angle is set according to these detected three parameters, the reference crank angle can be appropriately set according to the actual operating state of the internal combustion engine, The reference in-cylinder pressure and the actual in-cylinder pressure at the reference crank angle can be appropriately obtained. In addition, “detection” of various parameters in the present specification includes not only detecting the parameters directly with a sensor but also estimating them by calculation.

  According to a fourth aspect of the present invention, in the control device for an internal combustion engine according to any one of the first to third aspects, the crank angle at the start of compression at which compression of the air-fuel mixture is started in the compression stroke is set to an initial crank angle (intake air). Initial crank angle acquisition means (intake phase sensor 53, ECU 2) acquired as valve closing timing IVC) and initial in-cylinder temperature acquisition means (intake cylinder temperature T_STRT) that acquires the temperature in cylinder 3a at the start of compression as initial in-cylinder temperature T_STRT. The temperature sensor 57, the intake phase sensor 53, the exhaust phase sensor 54, the ECU 2, step 32 in FIG. 6, and the initial in-cylinder pressure acquisition that acquires the pressure in the cylinder 3a at the start of compression as the initial in-cylinder pressure (intake pressure PBA). Means (intake pressure sensor 56), and the reference in-cylinder pressure calculating means includes the reference crank angle CA_REF, the acquired initial crank angle, A period cylinder temperature T_STRT and initial cylinder pressure, according to the target EGR rate TGT_EGR, based on the temperature characteristics of the specific heat ratio of the gas mixture, to calculate the reference in-cylinder pressure P_REF (steps 33 in FIG. 6), wherein.

  The reference in-cylinder pressure is an in-cylinder pressure generated at a reference crank angle corresponding to immediately before the start of combustion of the air-fuel mixture. Therefore, the reference in-cylinder pressure changes according to the reference crank angle, and at the start of compression of the air-fuel mixture, It varies depending on the temperature and pressure of the gas mixture. Further, when the EGR rate is changed, the composition of the air-fuel mixture changes, and the specific heat ratio changes accordingly. Therefore, the reference in-cylinder pressure changes according to the EGR rate. From the above relationship, according to this configuration, the reference in-cylinder pressure is calculated according to the reference crank angle, the initial crank angle at the start of compression, the initial in-cylinder temperature and the initial in-cylinder pressure, and the target EGR rate. The in-cylinder pressure can be calculated appropriately.

  According to a fifth aspect of the present invention, in the control device for an internal combustion engine according to the fourth aspect, a rotational speed detection means (crank angle sensor 52) for detecting the rotational speed NE of the internal combustion engine, and cooling water for cooling the internal combustion engine 3 Cooling water temperature detecting means (water temperature sensor 59) for detecting the temperature TW of the engine, and the reference in-cylinder pressure calculating means includes a reference in-cylinder pressure according to the detected rotational speed NE of the internal combustion engine 3 and the cooling water temperature TW. It is characterized by correcting P_REF (steps 34 and 35 in FIG. 6).

  According to this configuration, by correcting the reference in-cylinder pressure according to the detected rotation speed of the internal combustion engine and the coolant temperature, it is possible to compensate for the influence of heat transferred between the cylinder and the outside. .

  According to a sixth aspect of the present invention, in the control device for an internal combustion engine according to the fourth or fifth aspect, the EGR rate estimating means is responsive to the reference crank angle CA_REF, the initial crank angle, the initial in-cylinder temperature T_STRT, and the initial in-cylinder pressure. Based on the temperature characteristic of the specific heat ratio of the air-fuel mixture, the EGR coefficient C_EGR representing the slope of the EGR rate R_EGR with respect to the pressure difference ΔP is calculated, and the target EGR is multiplied by the value obtained by multiplying the calculated EGR coefficient C_EGR by the pressure difference ΔP. It is characterized by calculating the EGR rate R_EGR by adding the rate TGT_EGR (steps 7, 11 and 13 in FIG. 3).

  The reference in-cylinder pressure is an in-cylinder pressure generated when the EGR rate is the target EGR rate. For this reason, when the actual EGR rate becomes equal to the target EGR rate, the actual in-cylinder pressure coincides with the reference in-cylinder pressure, the pressure difference between the two becomes 0, and the difference between the actual EGR rate and the target EGR rate is EGR. The greater the rate difference, the greater the pressure difference. Further, as shown in FIG. 5, the pressure difference has a linear relationship with the EGR rate (proportional relationship with the EGR rate difference), and the slope (proportional constant) varies depending on the intake conditions and the compression conditions. It recognized.

  From the above relationship, according to this configuration, when estimating the EGR rate, first, the EGR coefficient representing the slope of the EGR rate with respect to the pressure difference is set to the reference crank angle, the initial crank angle, the initial in-cylinder temperature, and the initial in-cylinder pressure. Calculate accordingly. As a result, among the parameters used for calculating the reference in-cylinder pressure, parameters other than the target EGR rate can be used to appropriately calculate the EGR coefficient while reflecting the intake / compression conditions of the air-fuel mixture. Moreover, since the EGR rate is calculated by adding the target EGR rate to the value obtained by multiplying the calculated EGR coefficient by the pressure difference, the EGR rate can be estimated with high accuracy.

  According to a seventh aspect of the present invention, in the control device for an internal combustion engine according to any one of the first to sixth aspects, the in-cylinder pressure sensor 51 is output from a pressure detection element for detecting the in-cylinder pressure, and the pressure detection element. The pressure detection element and the amplifier circuit are provided integrally with the fuel injection valve 4 that injects fuel into the cylinder 3a.

  The in-cylinder pressure sensor configured as described above has a pressure detection element and an amplifier circuit provided integrally with the fuel injection valve. Not easily affected. For this reason, the detection accuracy of the actual in-cylinder pressure by the in-cylinder pressure sensor is enhanced, and thus the estimation accuracy of the EGR rate can be further improved.

1 is a diagram schematically showing a configuration of an internal combustion engine to which the present invention is applied. It is a block diagram which shows schematic structure of a control apparatus. It is a main flow of the estimation process of an EGR rate. It is a subroutine which shows a reference crank angle setting process. It is a figure which shows the relationship between the pressure difference of an actual cylinder pressure-reference | standard cylinder pressure, an EGR rate, and a target EGR rate. It is a subroutine which shows the calculation process of the reference in-cylinder pressure. It is a figure which shows the input / output relationship of a reference | standard cylinder pressure map. It is a figure which shows the temperature characteristic of the specific heat ratio of each component of an air-fuel | gaseous mixture. It is a figure which shows the example of the relationship of the specific heat ratio of the air-fuel | gaseous mixture with respect to the crank angle in a compression stroke about the case where EGR gas does not exist and the case where it exists. It is a figure which shows the example of a setting of the reference in-cylinder pressure with respect to a reference | standard crank angle and an intake valve closing timing in a reference in-cylinder pressure map. It is a figure which shows the example of a setting of the reference in-cylinder pressure with respect to the initial in-cylinder temperature in a reference | standard in-cylinder pressure map. It is a figure which shows the example of a setting of the reference in-cylinder pressure with respect to intake pressure in a reference in-cylinder pressure map. It is a subroutine which shows the calculation process of an EGR coefficient. It is a figure which shows the input-output relationship of an EGR coefficient map. It is a figure which shows the example of a setting of the EGR coefficient with respect to a reference | standard crank angle and an intake valve closing timing in an EGR coefficient map. It is a figure which shows the example of a setting of the EGR coefficient with respect to the initial in-cylinder temperature in an EGR coefficient map. It is a figure which shows the example of a setting of the EGR coefficient with respect to intake pressure in an EGR coefficient map. It is a flowchart which shows the control processing of the ignition timing using an EGR rate. It is a subroutine which shows the calculation process of the reference in-cylinder pressure by a modification. It is a flowchart which shows the setting process of a target EGR rate. It is a figure which shows the example of a setting of the reference in-cylinder pressure with respect to the target EGR rate in a reference in-cylinder pressure map.

  Hereinafter, preferred embodiments of the present invention will be described in detail with reference to the drawings. As shown in FIG. 1, an internal combustion engine (hereinafter referred to as “engine”) 3 to which the present invention is applied is, for example, a gasoline engine having four cylinders 3a, and is mounted as a power source in a vehicle (not shown). .

  Each cylinder 3 a of the engine 3 is provided with a fuel injection valve (hereinafter referred to as “injector”) 4 and a spark plug 5 so as to face a combustion chamber (not shown) of the cylinder 3 a. The injector 4 is of a type that directly injects fuel into the combustion chamber. The spark discharge from the spark plug 5 ignites the fuel / air mixture and burns it. The fuel injection amount and fuel injection timing from the injector 4 and the ignition timing IGLOG of the spark plug 5 are controlled by a control signal from an electronic control unit (hereinafter referred to as “ECU”) 2 (see FIG. 2).

  In the present embodiment, the “air mixture” is an in-cylinder gas that is charged into the cylinder 3a and used for combustion. When exhaust gas recirculation (external EGR) is performed by the EGR device 14 described later, external EGR is performed. It contains gas.

  Each cylinder 3a of the engine 3 is provided with an in-cylinder pressure sensor 51 that detects an internal pressure (in-cylinder pressure). In this embodiment, the in-cylinder pressure sensor 51 is of an injector integrated type, and although not shown, faces the combustion chamber and amplifies and outputs a pressure detection element that picks up the in-cylinder pressure and a signal from the pressure detection element. An amplifier circuit or the like is integrally assembled with the injector 4. A detection signal representing the in-cylinder pressure PCYL detected by the in-cylinder pressure sensor 51 is input to the ECU 2.

  The engine 3 also includes a variable intake phase mechanism 11, a variable exhaust phase mechanism 12, a turbocharger 13, an EGR device 14, and the like.

  The variable intake phase mechanism 11 changes the relative phase (hereinafter referred to as “intake phase”) CAIN of an intake valve (both not shown) with respect to the crankshaft of the engine 3 in a stepless manner. 11a (see FIG. 2) and the like. The intake phase control motor 11a rotates the intake camshaft (not shown) relative to the crankshaft in response to a control signal from the ECU 2, and changes the relative angle between the two to change the intake phase CAIN. Change to stepless.

  Similarly, the variable exhaust phase mechanism 12 changes the relative phase (hereinafter referred to as “exhaust phase”) CAEX of an exhaust valve (not shown) with respect to the crankshaft continuously, and the exhaust phase control motor 12a ( Etc.). The exhaust phase control motor 12a rotates the exhaust camshaft (not shown) relative to the crankshaft in accordance with a control signal from the ECU 2, and changes the relative angle between the two to change the exhaust phase CAEX. Change to stepless.

  The variable intake phase mechanism 11 and the variable exhaust phase mechanism 12 control the timing of opening and closing the intake valve and the exhaust valve by changing the intake phase CAIN and the exhaust phase CAEX, respectively, and valve overflow between the intake valve and the exhaust valve. Used to control internal EGR by wrapping.

  The turbocharger 13 includes a compressor 21 provided in the intake passage 6 and a turbine 23 provided in the exhaust passage 7 and integrally connected to the compressor 21 via a shaft 22. The turbine 23 is driven by the exhaust gas flowing through the exhaust passage 7, and the compressor 21 rotates integrally therewith, whereby the intake air is supercharged. Further, the supercharging pressure is adjusted by controlling a waste gate valve (not shown) or the like with a control signal from the ECU 2.

  In the intake passage 6, an intake throttle valve 25, a compressor 21 of the turbocharger 13, an intercooler 26 for cooling intake air whose temperature has been increased by supercharging, and a throttle valve 27 are provided in this order from the upstream side. The intake throttle valve 25 generates a negative pressure for introducing the external EGR gas downstream thereof, and its opening degree is controlled via the LP actuator 25a in accordance with a control signal from the ECU 2.

  The throttle valve 27 is disposed upstream of the intake manifold 6 a in the intake passage 6. The opening degree of the throttle valve 27 is controlled via the TH actuator 27a in accordance with a control signal from the ECU 2, thereby controlling the in-cylinder gas amount sucked into the cylinder 3a.

  A three-way catalyst 28 is provided downstream of the turbine 23 in the exhaust passage 7. In the active state, the three-way catalyst 28 purifies the exhaust gas by oxidizing HC and CO in the exhaust gas and reducing NOx.

  The EGR device 14 recirculates a part of the exhaust gas discharged from the cylinder 3a into the exhaust passage 7 to the intake passage 6 as an external EGR gas via the EGR passage 41. As shown in FIG. 1, the EGR passage 41 is connected downstream of the turbine 23 and the three-way catalyst 28 in the exhaust passage 7 and between the compressor 21 and the intake throttle valve 25 in the intake passage 6. With this configuration, the external EGR gas has a relatively low pressure because the exhaust gas is taken out after the exhaust gas has worked on the turbine 23. That is, the EGR device 14 is configured as a so-called low pressure EGR device.

  In the middle of the EGR passage 41, an EGR valve 42 and an EGR cooler 43 for cooling the external EGR gas are provided. The opening degree of the EGR valve 42 is controlled via the EGR actuator 42a in accordance with a control signal from the ECU 2, thereby controlling the external EGR gas amount.

  In addition to the in-cylinder pressure sensor 51 described above, the engine 3 is provided with various sensors as described below in order to detect the operating state (see FIG. 2).

  The crank angle sensor 52 outputs a CRK signal and a TDC signal, which are pulse signals, to the ECU 2 for each predetermined crank angle as the crankshaft rotates. The CRK signal is output every predetermined crank angle (for example, 0.5 degrees). The ECU 2 calculates the engine speed (hereinafter referred to as “engine speed”) NE of the engine 3 based on the CRK signal.

  Further, the TDC signal is a signal indicating that the piston (not shown) of the engine 3 is in a predetermined crank angle position near the intake TDC (top dead center) in any of the cylinders 3a. Thus, when the engine 3 has four cylinders, it is output every crank angle of 180 degrees. The ECU 2 calculates a crank angle CA based on the output timing of the TDC signal for each cylinder 3a in accordance with the TDC signal and the CRK signal. Further, the ECU 2 calculates and assigns a crank angle stage FISTG (= 0 to 23) for each predetermined crank angle (for example, 30 degrees) according to the TDC signal and the CRK signal.

  The intake camshaft to which the variable intake phase mechanism 11 is attached and the exhaust camshaft to which the variable exhaust phase mechanism 12 is attached are provided with an intake phase sensor 53 and an exhaust phase sensor 54, respectively. The intake phase sensor 53 outputs a CAMIN signal, which is a pulse signal, to the ECU 2 at every predetermined cam angle (for example, 0.5 degrees) as the intake camshaft rotates. The ECU 2 calculates the intake phase CAIN based on the CAMIN signal and the CRK signal. Similarly, the exhaust phase sensor 54 outputs a CAMEX signal to the ECU 2 at every predetermined cam angle (for example, 0.5 degrees) as the exhaust camshaft rotates. The ECU 2 calculates the exhaust phase CAEX based on the CAMEX signal and the CRK signal.

  Further, in the intake passage 6, an air flow sensor 55 is provided on the upstream side of the intake throttle valve 25, and an intake pressure sensor 56 and an intake air temperature sensor 57 are provided in the intake chamber 6 b on the downstream side of the throttle valve 27. The air flow sensor 55 detects the amount (intake air amount) GAIR of air (fresh air) sucked into the cylinder 3a, the intake pressure sensor 56 detects the intake pressure PBA as an absolute pressure, and the intake air temperature sensor 57 An intake air temperature (intake air temperature) TA including the external EGR gas sucked into the cylinder 3a is detected. These detection signals are input to the ECU 2.

  A LAF sensor 58 is provided between the turbine 23 and the three-way catalyst 28 in the exhaust passage 7. The LAF sensor 58 continuously detects the oxygen concentration in the exhaust gas flowing into the three-way catalyst 28 in a wide air-fuel ratio region including the stoichiometric air-fuel ratio, and outputs a detection signal to the ECU 2. The ECU 2 calculates the exhaust gas equivalent ratio KACT based on this detection signal.

  Further, the ECU 2 receives a detection signal indicating the temperature TW of the cooling water for cooling the engine 3 from the water temperature sensor 59 (hereinafter referred to as “engine water temperature”) TW from the accelerator opening sensor 60 and an accelerator pedal (not shown) of the vehicle. Detection signals representing the amount of depression (hereinafter referred to as “accelerator opening”) AP are respectively input.

  The ECU 2 is composed of a microcomputer including a CPU, a RAM, a ROM, an I / O interface (all not shown), and the like. The ECU 2 determines the operating state of the engine 3 according to the detection signals of the various sensors described above, and executes engine control including control of the fuel injection amount of the injector 4 and the ignition timing IGLOG of the spark plug 5.

  Further, particularly in the present embodiment, the ECU 2 estimates the EGR rate R_EGR of the air-fuel mixture charged in the cylinder 3a, and controls the ignition timing IGLOG according to the estimated EGR rate R_EGR. The EGR rate R_EGR of the air-fuel mixture is defined as the ratio of the EGR gas amount to the total amount of the air-fuel mixture (cylinder gas).

  In the present embodiment, the ECU 2 corresponds to target EGR rate setting means, reference crank angle setting means, reference in-cylinder pressure calculation means, EGR rate estimation means, control means, initial crank angle acquisition means, and initial in-cylinder temperature acquisition means. .

  FIG. 20 shows the target EGR rate TGT_EGR setting process executed by the ECU 2. In this process, first, in step 61, a required torque TRQCMD is calculated by searching a predetermined map (not shown) according to the accelerator opening AP and the engine speed NE. Next, a target EGR rate TGT_EGR is calculated by searching a predetermined map (not shown) according to the required torque TRQCMD and the engine speed NE (step 62), and this process is terminated. When the target EGR rate TGT_EGR is set, a control signal corresponding to the target EGR rate TGT_EGR is output to the EGR actuator 42a, and the opening degree of the EGR valve 42 is controlled, whereby the EGR rate R_EGR is controlled to the target EGR rate TGT_EGR.

  Note that the above-described setting process of the target EGR rate TGT_EGR is merely an example, and other appropriate setting methods can be employed. For example, instead of the required torque TRQCMD, the target EGR rate TGT_EGR may be set using other parameters indicating the load of the engine 3, such as the intake air amount GAIR and the fuel injection amount, and further occurrence of knocking Corrections according to the situation may be added.

  Next, an estimation process for the EGR rate R_EGR executed by the ECU 2 will be described. FIG. 3 shows the main flow. This process is repeatedly executed for each cylinder 3a at the same cycle (for example, every 30 degrees of crank angle) as the switching cycle of the crank angle stage FISTG described above. In addition, the process directly related to the in-cylinder pressure PCYL detected by the in-cylinder pressure sensor 51 is executed at the same cycle as the CRK signal generation cycle (for example, every 0.5 degrees of crank angle) separately from the present process. For example, the detected in-cylinder pressure PCYL is stored corresponding to the crank angle CA.

  In the estimation process of FIG. 3, first, in step 1 (illustrated as “S1”, the same applies hereinafter), it is determined whether or not the crank angle stage FISTG is equal to a first predetermined value STG1 corresponding to intake TDC (top dead center). To do. When the determination result is YES and the cylinder 3a is in a stage immediately after the transition to the intake stroke, an intake-related parameter is acquired (step 2). Specifically, the intake air temperature TA, the engine water temperature TW, and the exhaust gas phase CAEX are read as intake-related parameters and stored in a predetermined area of the ECU 2 RAM. Thereafter, this process is terminated.

  If the determination result in step 1 is NO, it is determined whether or not the crank angle stage FISTG is equal to a second predetermined value STG2 corresponding to the compression BDC (bottom dead center) (step 3). When the determination result is YES and the cylinder 3a is in a stage immediately after the transition to the compression stroke, a compression related parameter is acquired (step 4). Specifically, the detected intake pressure PBA, engine speed NE, intake phase CAIN, and ignition timing IGLOG set at that time are read out as compression-related parameters and stored in a predetermined area of the RAM of the ECU 2. .

  Next, a reference crank angle CA_REF setting process is executed (step 5). This setting process predicts the timing immediately before the start of combustion of the air-fuel mixture and sets it as the reference crank angle CA_REF. FIG. 4 shows the subroutine.

  In this process, first, in step 21, a retardation correction amount ΔC_CA is calculated by searching a predetermined map (not shown) according to the intake pressure PBA and the engine speed NE acquired in step 4. This retard correction amount ΔC_CA corresponds to the ignition delay time from when the ignition operation is performed by the spark plug 5 at the ignition timing IGLOG to when the air-fuel mixture is ignited and combustion is started, and is represented by the crank angle. . The lower the intake pressure PBA, the more difficult it is for the air-fuel mixture to ignite, and the higher the engine speed NE, the larger the crank angle corresponding to the same ignition delay time. Therefore, in the above map, the retardation correction amount ΔC_CA is set to be larger as the intake pressure PBA is lower and the engine speed NE is higher.

  Next, a value obtained by subtracting the retardation correction amount ΔC_CA from the ignition timing IGLOG acquired in step 4 is set as a reference crank angle CA_REF (step 22), and this process is terminated. The reference crank angle CA_REF is expressed with the compression TDC of each cylinder 3a as the origin (0 degree) and the advance side as positive (see FIG. 10).

  Returning to FIG. 3, in step 6 following step 5, the reference in-cylinder pressure P_REF is calculated. The reference in-cylinder pressure P_REF is equal to the reference crank angle under the condition that the EGR rate R_EGR of the air-fuel mixture is equal to the target EGR rate TGT_EGR set in the processing of FIG. 20 and the air-fuel ratio of the air-fuel mixture is the stoichiometric air-fuel ratio. This is the in-cylinder pressure generated at. Details of the calculation process will be described later.

  Next, an EGR coefficient C_EGR calculation process is executed (step 7), and this process ends. As shown in FIG. 5, the EGR coefficient C_EGR is a difference between a pressure difference ΔP (a difference between an actual in-cylinder pressure P_CPS and a reference in-cylinder pressure P_REF described later) and an EGR rate difference ΔEGR (a difference between the actual EGR rate R_EGR and the target EGR rate TGT_EGR ), A proportional relationship is recognized, and the slope (= ΔEGR / ΔP) is defined as the EGR coefficient C_EGR. Details of the calculation process will be described later.

  When the determination result in step 3 is NO, it is determined whether or not the crank angle stage FISTG is equal to a third predetermined value STG3 corresponding to compression TDC (top dead center) (step 8). When this determination result is NO, this process is terminated as it is. On the other hand, when the determination result in step 8 is YES and the cylinder 3a is in a stage immediately after the compression stroke is finished, the in-cylinder pressure PCYL detected at the reference crank angle CA_REF set in step 5 is read from the RAM, Acquired as in-cylinder pressure P_CPS (step 9).

Next, a difference (= P_CPS−P_REF) between the acquired actual in-cylinder pressure P_CPS and the reference in-cylinder pressure P_REF calculated in step 6 is calculated as a pressure difference ΔP (step 10). Next, using the pressure difference ΔP and EGR coefficient C_EGR calculated so far and the target EGR rate TGT_EGR, the EGR rate R_EGR of the air-fuel mixture is calculated according to the following equation (A) (step 11), and this process ends.
R_EGR = ΔP × C_EGR + TGT_EGR (A)
The equation (A) is directly derived from the following equation (A) ′ that represents the definition of the EGR coefficient C_EGR described above (see FIG. 5).
C_EGR = ΔEGR / ΔP
= (R_EGR−TGT_EGR) / ΔP (A) ′

  Next, the reference in-cylinder pressure P_REF calculation process executed in step 6 of FIG. 3 will be described with reference to FIG. In this process, first, in step 31, the intake valve closing timing (hereinafter referred to as “intake valve closing timing”) IVC is calculated from the intake phase CAIN acquired in step 2. The intake valve closing timing IVC is represented by a crank angle in which the compression TDC is the origin (0 degree) and the advance side is positive, like the reference crank angle CA_REF described above.

  When the intake valve closing timing IVC is set during the compression stroke, the compression of the air-fuel mixture is substantially started when the intake valve is closed. Therefore, the intake valve closing timing IVC is determined by the crank angle ( This corresponds to the initial crank angle. The intake pressure PBA corresponds to the in-cylinder pressure (initial in-cylinder pressure) at the start of compression.

  Next, by searching a predetermined map (not shown) according to the intake air temperature TA, the intake air phase CAIN, and the exhaust gas phase CAEX, an initial in-cylinder temperature T_STRT that is the temperature in the cylinder 3a at the start of compression is calculated. (Step 32). Among the above parameters, the intake phase CAIN and the exhaust phase CAEX reflect an increase in the in-cylinder temperature corresponding to the internal EGR amount when the internal EGR due to the valve overlap between the intake valve and the exhaust valve is executed. belongs to. Therefore, in the above map, the initial in-cylinder temperature T_STRT has a higher value as the intake air temperature TA is higher and as the intake valve phase CAIN and the exhaust gas phase CAEX are closer to the valve overlap side. Is set to

  In the next step 33, the reference in-cylinder pressure map shown in FIG. 7 is retrieved by searching the reference in-cylinder pressure map shown in FIG. 7 according to the reference crank angle CA_REF, the intake valve closing timing IVC, the initial in-cylinder temperature T_STRT, the intake pressure PBA, and the target EGR rate TGT_EGR. In-cylinder pressure P_REF is calculated. Hereinafter, the reference in-cylinder pressure map will be described.

First, the specific heat ratio of the air-fuel mixture (cylinder gas) filled in the cylinder 3a and the state change in the compression stroke will be described. The specific heat ratio κ of the air-fuel mixture is expressed by the following equation (1) using the constant pressure specific heat Cp and the gas constant R, and the constant pressure specific heat Cp is expressed by the following equation (2).

  As shown in Expression (2), the specific heat ratio κ of the air-fuel mixture varies depending on its composition (components and the number of moles of each component). Further, as shown in FIG. 8, the specific heat ratio of each component of the air-fuel mixture has a temperature characteristic of decreasing as the temperature rises, and the specific heat ratio κ of the air-fuel mixture composed of these components is also the same. Temperature characteristics. Furthermore, as shown in FIG. 9, when the mixture contains EGR gas, the composition of the mixture changes, and the CO2 component of the EGR gas is added to increase the specific heat ratio κ of the mixture. Has characteristics.

In addition, since the state change of the air-fuel mixture in the compression stroke is an adiabatic compression change and is regarded as a polytropic change, the in-cylinder temperature Ta when the crank angle CA = a is expressed by the following equation (3).

As shown in Expression (3), the in-cylinder temperature T is a function of the specific heat ratio κ, and the specific heat ratio κ of the air-fuel mixture is a function of the in-cylinder temperature T as described above. For this reason, in order to accurately obtain the specific heat ratio κ and the in-cylinder temperature T, sequential calculation is performed by sequentially applying the calculation results of the expressions (1), (2), and (3) to each other. As a result, the in-cylinder temperature (final in-cylinder temperature) Tθ when the crank angle CA = the final crank angle θ is expressed by the following equation (4).

Further, the in-cylinder pressure Pa when the crank angle CA = a is expressed by the following equation (5). Further, from this equation (5), the in-cylinder pressure (final in-cylinder pressure) Pθ when the crank angle CA = θ is Is represented by the following equation (6).

  As shown in the equation (6), the final in-cylinder pressure Pθ is a function of the initial in-cylinder pressure P0, the initial cylinder volume V0, the final cylinder volume Vθ, and the specific heat ratio κ that is sequentially calculated. Further, the specific heat ratio κ is a function of the in-cylinder temperature T calculated sequentially, and the in-cylinder temperature T is a function of the initial in-cylinder temperature T0 and the specific heat ratio κ. Further, since the cylinder volume V is uniquely obtained from the crank angle CA, the initial cylinder volume V0 and the final cylinder volume Vθ are replaced with the initial crank angle CA0 and the final crank angle CAθ, respectively.

  From the above, the final in-cylinder pressure Pθ is obtained as a function of the initial in-cylinder pressure P0, the initial in-cylinder temperature T0, the initial crank angle CA0, and the final crank angle CAθ under the condition given the composition of the air-fuel mixture in the equation (2). It is done.

  The above-described reference in-cylinder pressure map is based on the above relationship. As shown in FIG. 7, the in-cylinder pressure PBA and the initial in-cylinder temperature corresponding to the initial in-cylinder pressure P0, the initial in-cylinder temperature T0, and the initial crank angle CA0, respectively. T_STRT and intake valve closing timing IVC and a reference crank angle CA_REF corresponding to the final crank angle CAθ are used as input parameters, and a reference cylinder pressure P_REF corresponding to the final cylinder pressure Pθ is obtained as an output.

  Further, the target EGR rate TGT_EGR set at that time is added to the input parameter as a condition for the composition of the air-fuel mixture. From this target EGR rate TGT_EGR and the condition that the air-fuel ratio is the stoichiometric air-fuel ratio, the equation (2 The number of moles nx of each component is determined. Specifically, the reference in-cylinder pressure P_REF is calculated in advance based on the formulas (1) to (6) for various conditions of the above five input parameters, and the calculation result is mapped to the input parameters. Thus, a reference in-cylinder pressure map is created.

  10 to 12 and 21 show examples of setting the reference in-cylinder pressure P_REF for each input parameter in the reference in-cylinder pressure map. First, as shown in FIG. 21, the reference in-cylinder pressure P_REF is set to a larger value as the target EGR rate C_EGR is higher. This is because the higher the EGR rate, the higher the specific heat ratio κ of the air-fuel mixture (see FIG. 9), and the final in-cylinder pressure increases accordingly.

  Further, as shown in FIG. 10, the reference in-cylinder pressure P_REF is set to a larger value as the value of the reference crank angle CA_REF is closer to 0, that is, as the reference crank angle CA_REF is closer to the compression TDC. The larger the value of the valve timing IVC, that is, the earlier the closing timing of the intake valve in the compression stroke, the larger the value is set. This is because, as the reference crank angle CA_REF is closer to the compression TDC and the closing timing of the intake valve is earlier, the substantial compression period of the air-fuel mixture becomes longer, and the final in-cylinder pressure increases. is there.

  As shown in FIG. 11, the reference in-cylinder pressure P_REF is set to a smaller value as the initial in-cylinder temperature T_STRT is higher. This is because as the initial in-cylinder temperature T_STRT is higher, the specific heat ratio κ of the air-fuel mixture decreases as the in-cylinder temperature becomes higher, and as a result, the increase in the in-cylinder pressure decreases.

  Further, as shown in FIG. 12, the reference in-cylinder pressure P_REF is set to be proportional to the intake pressure PBA. This is because the reference in-cylinder pressure P_REF and the intake pressure PBA correspond to the final in-cylinder pressure Pθ and the initial in-cylinder pressure P0, respectively, and are in a proportional relationship (see Expression (6)).

  As described above, in step 33 of FIG. 6, the reference in-cylinder pressure P_REF is calculated by searching the reference in-cylinder pressure map described above according to the above five parameters. In the next step 34, a heat transfer correction coefficient K_HT is calculated by searching a predetermined map in accordance with the engine speed NE and the engine water temperature TW. The heat transfer correction coefficient K_HT is for compensating for the influence of heat exchanged between the cylinder 3a and the outside.

  Next, the final reference in-cylinder pressure P_REF is calculated by multiplying the reference in-cylinder pressure P_REF calculated in step 33 by the heat transfer correction coefficient K_HT (step 35), and this process is terminated.

  Next, the EGR coefficient C_EGR calculation process executed in step 7 of FIG. 3 will be described with reference to FIG. As described above, the EGR coefficient C_EGR is defined as the slope of the EGR rate difference ΔEGR (the difference between the actual EGR rate R_EGR and the target EGR rate TGT_EGR) with respect to the pressure difference ΔP (the difference between the actual in-cylinder pressure P_CPS and the reference in-cylinder pressure P_REF). (See FIG. 5) and used to calculate the EGR rate R_EGR. In addition, since the characteristic that the inclination changes according to the intake condition and the compression condition is recognized, the EGR coefficient C_EGR is calculated in this process.

  In this process, first, in step 41, the reference crank angle CA_REF, the intake valve closing timing IVC, the initial in-cylinder temperature T_STRT, and the intake pressure PBA are acquired. These parameters represent the intake conditions and the compression conditions described above, and are common to the four input parameters of the reference in-cylinder pressure map described above. For this reason, the parameter acquisition in step 41 is performed by reading the data obtained by the calculation process of the reference in-cylinder pressure P_REF in FIG.

  Next, an EGR coefficient C_EGR is calculated by searching the EGR coefficient map shown in FIG. 14 according to the acquired four parameters (step 42), and this process is terminated. This EGR coefficient map is obtained by previously calculating the EGR coefficient C_EGR based on the equations (1) to (6) for various conditions of the above four input parameters, and mapping the result to the input parameters. It is. Unlike the reference in-cylinder pressure P_REF, the EGR coefficient C_EGR is not affected by the target EGR rate TGT_EGR, and therefore the target EGR rate TGT_EGR is not included in the input parameters of the EGR coefficient map.

  15 to 17 show setting examples of the EGR coefficient C_EGR for each input parameter in the EGR coefficient map. As shown in FIG. 15, the EGR coefficient C_EGR is set to a smaller value as the reference crank angle CA_REF is closer to the compression TDC and as the closing timing of the intake valve in the compression stroke is earlier. This is because, as the reference crank angle CA_REF is closer to the compression TDC and the closing timing of the intake valve is earlier, the substantial compression period of the air-fuel mixture becomes longer, and the pressure difference ΔP increases accordingly. This is because the EGR coefficient C_EGR becomes small.

  As shown in FIG. 16, the EGR coefficient C_EGR is set to a smaller value as the initial in-cylinder temperature T_STRT is higher. This is due to the following reason. That is, among the components of the air-fuel mixture, the fuel has a large temperature change in the constant pressure specific heat Cp and the contribution of the specific heat ratio κ of the air-fuel mixture to the temperature characteristics compared to other components. On the other hand, when the EGR rate R_EGR increases, the proportion of fuel decreases accordingly, and the degree of change in the specific heat ratio κ according to the temperature decreases as the contribution decreases. For this reason, as the initial in-cylinder temperature T_STRT is higher, the specific heat ratio κ is changed at a higher level during compression, so that the pressure difference ΔP increases and the EGR coefficient C_EGR decreases.

  Further, as shown in FIG. 17, the EGR coefficient C_EGR is set to a smaller value as the intake pressure PBA is higher. This is because the actual in-cylinder pressure P_CPS and the pressure difference ΔP increase in proportion to the intake pressure PBA that is the initial in-cylinder pressure, and the EGR coefficient C_EGR decreases accordingly.

  Here, referring to FIG. 5 again, an operation obtained by the above-described estimation processing of the EGR rate R_EGR will be described. For example, when the target EGR rate TGT_EGR shown in FIG. 5 is set from the state where the EGR is stopped (target EGR rate TGT_EGR = 0) and the external EGR is started, a significant delay time is given to the low-pressure EGR device from the start. Until the actual EGR rate R_EGR is 0, the EGR rate difference ΔEGR is a negative value, and the actual in-cylinder pressure P_CPS is smaller than the reference in-cylinder pressure PREF, so that the pressure difference ΔP is also negative. Value.

  When the above delay time elapses and the external EGR gas is introduced into the cylinder 3a, the EGR rate R_EGR increases from 0 and finally reaches the target EGR rate TGT_EGR. As a result, the actual in-cylinder pressure P_CPS increases, so that the pressure difference ΔP increases linearly with a slope corresponding to the EGR coefficient C_EGR, as shown by an arrow A in the figure, and the EGR rate R_EGR becomes the target EGR rate TGT_EGR. When it reaches, it becomes 0. Therefore, when the pressure difference ΔP = ΔP1 is obtained at any point in the meantime, the actual EGR rate R_EGR1 at that time can be accurately calculated by the above equation (A) using the EGR coefficient C_EGR and the target EGR rate TGT_EGR. Can do.

  On the contrary, even when the target EGR rate TGT_EGR is changed from a larger value to a smaller value, the EGR rate R_EGR can be calculated in the same manner as described above. That is, in this case, the actual EGR rate R_EGR remains the target EGR rate before the change until the delay time elapses from the start of the change of the target EGR rate TGT_EGR, so that the EGR rate difference ΔEGR is a positive value. Accordingly, the pressure difference ΔP is also a positive value.

  When the delay time elapses and the amount of external EGR gas introduced into the cylinder 3a decreases, the EGR rate R_EGR decreases and finally reaches the target EGR rate TGT_EGR. Accordingly, as the actual in-cylinder pressure P_CPS decreases, the pressure difference ΔP decreases linearly with a slope corresponding to the EGR coefficient C_EGR, as indicated by an arrow B in the figure, and EGR rate R_EGR = target EGR rate TGT_EGR When it becomes, it becomes 0. Therefore, when the pressure difference ΔP = ΔP2 is obtained at an arbitrary point in the meantime, the actual EGR rate R_EGR2 at that time can be accurately calculated by the equation (A) as in the case of ΔP1.

  Next, the ignition timing control process using the EGR rate R_EGR will be described with reference to FIG. This process is executed for each cylinder 3a in synchronization with the generation of the TDC signal. In this process, first, in step 51, a basic ignition timing IG_BASE is calculated by searching a predetermined map (not shown) according to the engine speed NE and the required torque TRQCMD.

  Next, an EGR correction amount ΔIGEGR is calculated by searching a predetermined map (not shown) according to the estimated EGR rate R_EGR (step 52).

  Next, a correction amount ΔIGTTL due to factors other than the EGR rate R_EGR is calculated according to the engine water temperature TW, the engine speed NE, and the like (step 53).

  Finally, the ignition timing IGLOG is calculated by adding the EGR correction amount ΔIGEGR and the correction amount ΔIGTTL to the basic ignition timing IG_BASE (step 54), and this processing is terminated.

  As described above, according to the present embodiment, the temperature characteristic of the specific heat ratio κ of the mixture is obtained under the condition of the mixture composition in which the EGR rate R_EGR is equal to the target EGR rate TGT_EGR and the air-fuel ratio is the stoichiometric air-fuel ratio. Based on this, the reference in-cylinder pressure P_REF generated at the reference crank angle CA_REF is calculated. And, since the EGR rate R_EGR is calculated based on the pressure difference ΔP between the actual in-cylinder pressure P_CPS detected at the reference crank angle CA_REF and the reference in-cylinder pressure P_REF, while reflecting the temperature characteristic of the specific heat ratio κ of the air-fuel mixture, The EGR rate R_EGR can be estimated.

  Further, since the reference crank angle CA_REF is the crank angle immediately before the start of combustion of the air-fuel mixture, the actual in-cylinder pressure P_CPS is acquired in a state where combustion is not yet performed and the state change of the air-fuel mixture is maintained at the polytropic change. In addition, a large pressure difference ΔP between the actual in-cylinder pressure P_CPS and the reference in-cylinder pressure P_REF can be secured. Therefore, based on this pressure difference ΔP, the EGR rate EGR rate R_EGR can be accurately estimated while favorably reflecting the temperature characteristics of the specific heat ratio κ of the air-fuel mixture. Further, the ignition timing IGLOG can be appropriately controlled using the EGR rate R_EGR estimated with high accuracy.

  In addition, since the estimation of the EGR rate R_EGR is performed using the actual in-cylinder pressure P_CPS and the reference in-cylinder pressure P_REF as the actual pressure in the cylinder 3a as parameters, even when the EGR device 14 is a low-pressure EGR device, The EGR rate R_EGR can be accurately estimated without being affected by the delay.

  Further, since the reference crank angle CA_REF is set using the ignition timing IGLOG, the intake pressure PBA, and the engine speed NE, the setting can be appropriately performed according to the actual operating state of the engine 3, and the reference crank angle CA_REF can be set appropriately. The reference in-cylinder pressure P_REF and the actual in-cylinder pressure P_CPS at the angle CA_REF can be appropriately obtained.

  Further, the reference in-cylinder pressure P_REF is set to the reference crank angle CA_REF, the intake valve closing timing IVC corresponding to the initial crank angle at the start of compression, the initial in-cylinder temperature T_STRT, the intake pressure PBA corresponding to the initial in-cylinder pressure, and the target EGR rate. It is possible to calculate appropriately according to TGT_EGR. Further, by correcting the calculated reference in-cylinder pressure P_REF according to the engine speed NE and the engine water temperature TW, it is possible to compensate for the influence of heat transferred between the inside of the cylinder 3a and the outside.

  Further, among the parameters used for calculating the reference in-cylinder pressure P_REF, the four parameters other than the target EGR rate TGT_EGR (reference crank angle CA_REF, intake valve closing timing IVC, initial in-cylinder temperature T_STRT, and intake pressure PBA) are mixed. The EGR coefficient C_EGR can be calculated appropriately while reflecting the intake / compression conditions of the air. Then, the EGR rate R_EGR is calculated by multiplying the calculated EGR coefficient C_EGR by the pressure difference ΔP and further adding the target EGR rate TGT_EGR. Therefore, the EGR rate R_EGR can be estimated with high accuracy.

  Furthermore, since the pressure detecting element and the amplifier circuit are integrally provided in the injector 4, the in-cylinder pressure sensor 51 is not easily affected by noise due to ignition operation or noise due to the injection operation of the injector 4 of another cylinder 3 a. For this reason, the detection accuracy of the actual in-cylinder pressure P_CPS by the in-cylinder pressure sensor 51 is increased, so that the estimation accuracy of the EGR rate R_EGR can be further improved.

  Next, a modified example of the calculation process of the reference in-cylinder pressure P_REF will be described with reference to FIG. As described above, in this modification, the reference in-cylinder pressure P_REF is proportional to the intake pressure PBA (FIG. 12), and therefore, the intake pressure PBA is excluded from the input parameters of the reference in-cylinder pressure map, and is obtained in the reference in-cylinder pressure map. The obtained map value is corrected by the intake pressure PBA. This process is executed instead of the process of FIG. In FIG. 19, steps having the same processing contents as those in FIG. 6 are given the same step numbers.

  In this process, the same step 31 and step 32 as the process of FIG. 6 are executed to calculate the intake valve closing timing IVC and the initial in-cylinder temperature T_STRT. Next, a reference in-cylinder pressure P_REF is calculated by searching a reference in-cylinder pressure map (not shown) according to the reference crank angle CA_REF, the intake valve closing timing IVC, the initial in-cylinder temperature T_STRT, and the target EGR rate TGT_EGR ( Step 301). In this reference in-cylinder pressure map, the initial in-cylinder pressure at the start of compression is treated as a constant, and the reference atmospheric pressure PATM (760 mmHg) is used.

  Next, a value obtained by dividing the intake pressure PBA by the reference atmospheric pressure PATM is set as an intake pressure correction coefficient K_PB (step 302), and this intake pressure correction coefficient K_PB is multiplied by the reference in-cylinder pressure P_REF calculated in step 301. Thus, the corrected reference in-cylinder pressure P_REF is calculated (step 303).

  The subsequent processing contents are the same as in FIG. 6, and the final in-cylinder pressure P_REF calculated in step 303 is multiplied by the heat transfer correction coefficient K_HT calculated according to the engine speed NE and the engine water temperature TW. A basic reference in-cylinder pressure P_REF is calculated (steps 34 and 35), and this process is terminated.

  According to the above modification, the reference in-cylinder pressure P_REF equivalent to that in the calculation process of FIG. 6 can be calculated, and the input parameter is reduced, so that the reference in-cylinder pressure map can be easily created and the load is reduced. can do.

  In addition, this invention can be implemented in various aspects, without being limited to the described embodiment. For example, in the embodiment, the ignition timing IGLOG, the intake pressure PBA, and the engine speed NE are used as parameters when calculating the reference crank angle CA_REF. However, other appropriate parameters may be used in combination.

  In the embodiment, the initial in-cylinder temperature T_STRT used for calculating the reference in-cylinder pressure P_REF and the EGR coefficient C_EGR is calculated according to the intake air temperature TA, the intake phase CAIN, and the exhaust phase CAEX. When the internal EGR due to the overlap is not executed, the intake air temperature TA may be used as the initial in-cylinder temperature as it is. Further, although the intake pressure PBA is used as the initial in-cylinder pressure, it is also possible to use the in-cylinder pressure PCYL detected by the in-cylinder pressure sensor 51 at the start of compression.

  Further, the reference in-cylinder pressure P_REF is corrected according to the engine speed NE and the engine water temperature TW, but the correction is further performed using other appropriate parameters that affect the transfer of heat between the inside and outside of the cylinder 3a. You may go.

  Furthermore, in the embodiment, the ignition timing control is executed in accordance with the estimated EGR rate R_EGR. However, instead of or in addition to this, other engine control, for example, EGR control via the EGR valve 42, throttle valve The intake air amount control via 27 or the fuel injection control via the injector 4 may be executed.

  In the embodiment, the EGR device 14 is configured by a low-pressure EGR device. However, instead of or in addition to this, a high-pressure EGR device may be used, and in this case, the above-described effects can be obtained similarly. Can do. Furthermore, although the in-cylinder pressure sensor 51 is integrated with the injector 4, it is needless to say that the in-cylinder pressure sensor 51 may be a separate type arranged separately from the injector 4.

  Furthermore, in the embodiment, the engine 3 is a vehicle engine, but the present invention can also be applied to an engine for other uses, for example, an engine for an outboard motor in which a crankshaft is arranged in the vertical direction. In addition, the detailed configuration can be changed as appropriate within the scope of the gist of the present invention.

2 ECU (target EGR rate setting means, reference crank angle setting means, reference in-cylinder pressure calculation means, EGR rate estimation means, control means, initial crank angle acquisition means, initial in-cylinder temperature acquisition means)
3 Internal combustion engine 3a Cylinder 4 Fuel injection valve 6 Intake passage 7 Exhaust passage 13 Turbocharger (supercharger)
14 EGR device 21 Compressor 23 Turbine 51 In-cylinder pressure sensor 52 Crank angle sensor (operation state detection means, rotation speed detection means)
53 Intake phase sensor (initial crank angle acquisition means, initial in-cylinder temperature acquisition means)
54 Exhaust phase sensor (initial cylinder temperature acquisition means)
56 Intake pressure sensor (operating state detection means, initial in-cylinder pressure acquisition means)
57 Intake air temperature sensor (initial cylinder temperature acquisition means)
59 Water temperature sensor (cooling water temperature detection means)
κ Specific heat ratio of gas mixture TGT_EGR Target EGR rate PCYL In-cylinder pressure
CA crank angle CA_REF reference crank angle P_REF reference cylinder pressure P_CPS actual cylinder pressure
ΔP Pressure difference between actual in-cylinder pressure and reference in-cylinder pressure R_EGR EGR rate IGLOG Ignition timing PBA Intake pressure (intake pressure, initial in-cylinder pressure)
NE engine speed (speed of internal combustion engine)
IVC Intake valve closing timing (initial crank angle)
T_STRT Initial in-cylinder temperature
TW engine water temperature (cooling water temperature)
C_EGR EGR coefficient

Claims (7)

A control device for an internal combustion engine comprising an EGR device that directly injects fuel into a cylinder and recirculates a part of exhaust gas discharged from the cylinder to an exhaust passage as an external EGR gas to an intake passage,
Target EGR rate setting means for setting a target EGR rate which is a target of the EGR rate of the air-fuel mixture charged in the cylinder;
An in-cylinder pressure sensor for detecting a pressure in the cylinder as an in-cylinder pressure;
An operating state detecting means for detecting an operating state of the internal combustion engine;
Reference crank angle setting means for setting, as a reference crank angle, a crank angle immediately before the combustion of the air-fuel mixture is started according to the detected operating state of the internal combustion engine;
Based on the temperature characteristics of the specific heat ratio of the mixture, the EGR rate of the mixture is equal to the set target EGR rate and the air-fuel ratio of the mixture is the stoichiometric air-fuel ratio. A reference in-cylinder pressure calculating means for calculating a generated in-cylinder pressure as a reference in-cylinder pressure;
EGR rate estimating means for estimating the EGR rate of the air-fuel mixture based on the pressure difference between the actual in-cylinder pressure detected by the in-cylinder pressure sensor at the reference crank angle and the calculated reference in-cylinder pressure;
Control means for controlling the internal combustion engine in accordance with the estimated EGR rate;
A control device for an internal combustion engine, comprising:   The EGR device is configured to recirculate external EGR gas from a downstream side of a turbocharger turbine in the exhaust passage to an upstream side of the turbocharger compressor in the intake passage. The control device for an internal combustion engine according to claim 1. The operating state detecting means detects an ignition timing, a pressure of intake air sucked into the cylinder, and a rotational speed of the internal combustion engine as an operating state of the internal combustion engine,
The internal combustion engine according to claim 1 or 2, wherein the reference crank angle setting means sets the reference crank angle in accordance with the detected ignition timing, intake pressure and engine speed. Control device. Initial crank angle obtaining means for obtaining a crank angle at the start of compression at which compression of the air-fuel mixture is started in the compression stroke as an initial crank angle;
Initial in-cylinder temperature acquisition means for acquiring the temperature in the cylinder at the start of compression as an initial in-cylinder temperature;
An initial in-cylinder pressure acquiring means for acquiring the pressure in the cylinder at the start of compression as an initial in-cylinder pressure;
The reference in-cylinder pressure calculating means is characterized in that the reference crank angle, the acquired initial crank angle, the initial in-cylinder temperature and the initial in-cylinder pressure, and the temperature characteristic of the specific heat ratio of the air-fuel mixture according to the set target EGR rate. The control apparatus for an internal combustion engine according to any one of claims 1 to 3, wherein the reference in-cylinder pressure is calculated based on the equation (1). A rotational speed detection means for detecting the rotational speed of the internal combustion engine;
Cooling water temperature detecting means for detecting the temperature of cooling water for cooling the internal combustion engine,
5. The control device for an internal combustion engine according to claim 4, wherein the reference in-cylinder pressure calculating unit corrects the reference in-cylinder pressure in accordance with the detected rotational speed of the internal combustion engine and the temperature of the cooling water. .   The EGR rate estimating means is configured to determine the EGR rate with respect to the pressure difference based on a temperature characteristic of a specific heat ratio of a mixture according to the reference crank angle, the initial crank angle, the initial in-cylinder temperature, and the initial in-cylinder pressure. And calculating the EGR rate by adding the target EGR rate to a value obtained by multiplying the pressure difference by the calculated EGR coefficient. Item 6. The control device for an internal combustion engine according to Item 4 or 5.   The in-cylinder pressure sensor includes a pressure detection element for detecting the in-cylinder pressure, and an amplification circuit that amplifies and outputs a signal output from the pressure detection element. 7. The control device for an internal combustion engine according to claim 1, wherein the control device is provided integrally with a fuel injection valve for injecting fuel into the cylinder.

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