Control device for internal combustion engine

A control device includes a cylinder pressure sensor, a driving condition detector, a reference crank angle setter, a reference cylinder pressure calculator, an EGR ratio estimator, and a controller. The cylinder pressure sensor detects a cylinder pressure inside a cylinder. The driving condition detector detects a driving condition in an internal combustion engine. The reference crank angle setter calculates, according to the driving condition, a reference crank angle immediately before which mixture gas starts combusting. The reference cylinder pressure calculator calculates a reference cylinder pressure at the reference crank angle based on temperature characteristics of a heat capacity ratio of the mixture gas under a condition. The EGR ratio estimator calculates an EGR ratio based on a pressure difference between the reference cylinder pressure and the cylinder pressure at the reference crank angle. The controller controls the internal combustion engine according to the EGR ratio.

CROSS-REFERENCE TO RELATED APPLICATIONS

The present application claims priority under 35 U.S.C. § 119 to Japanese Patent Application No. 2015-138597, filed Jul. 10, 2015, entitled “Control Device for Internal Combustion Engine” and Japanese Patent Application No. 2015-241446, filed Dec. 10, 2015, entitled “Control Device for Internal Combustion Engine.” The contents of these applications are incorporated herein by reference in their entirety.

BACKGROUND

The present disclosure relates to a control device for an internal combustion engine.

2. Description of the Related Art

There has been known a conventional EGR ratio estimating method which involves detecting an amount of fresh air to be taken into a cylinder by an airflow sensor, calculating a total amount of gas to be taken into the cylinder based on an intake pressure detected by an intake pressure sensor, and estimating an EGR ratio from the amount of fresh air and the total amount of gas. In the case of using this estimating method for a low-pressure EGR device (EGR device which takes in an exhaust gas from a downstream side of a turbine of a turbocharger and recirculates the exhaust gas back to an upstream side of a compressor in an intake passage), however, it is difficult to estimate the EGR ratio with high accuracy because the external EGR gas needs to flow a relatively long passage leading to the cylinder, and accordingly reaches the inside of the cylinder with a long time lag.

As another conventional EGR ratio estimating method, a method disclosed in Japanese Unexamined Patent Application Publication No. 2008-231995 has been known, for example. This estimating method is based on the assumption that a state change of a mixture gas in a compression stroke of the internal combustion engine is a polytropic change, and that a heat capacity ratio of the mixture gas varies depending on its composition. More specifically, an in-cylinder pressure sensor detects in-cylinder pressures P1, P2at two predetermined crank angles CA1, CA2in the compression stroke, and a heat capacity ratio κ of the mixture gas is calculated from the in-cylinder pressures P1, P2and cylinder volumes V1, V2corresponding to the crank angles CA1, CA2in accordance with the following formula:
κ=log(P1/P2)/log(V2/V1).

Then, the EGR ratio is calculated based on the calculated heat capacity ratio κ by referring to a predetermined table in which a relationship between the heat capacity ratio and the EGR ratio (EGR gas concentration) is defined.

SUMMARY

According to one aspect of the present invention, a control device for an internal combustion engine including an EGR device that, concurrently with direct injection of a fuel into a cylinder, recirculates a portion of an exhaust gas discharged to an exhaust passage from the cylinder back to an intake passage as an external EGR gas, the control device includes an in-cylinder pressure sensor, a driving condition detector, a reference crank angle setter, a reference in-cylinder pressure calculator, an EGR ratio estimator, and a controller. The in-cylinder pressure sensor detects a pressure inside the cylinder as an in-cylinder pressure. The driving condition detector detects a driving condition of the internal combustion engine. The reference crank angle setter obtains a crank angle immediately before start of combustion of a mixture gas charged in the cylinder depending on the detected driving condition of the internal combustion engine, and sets the obtained crank angle as a reference crank angle. The reference in-cylinder pressure calculator calculates, as a reference in-cylinder pressure, a pressure inside the cylinder expected to occur at the set reference crank angle, based on a temperature property of a heat capacity ratio of the mixture gas under conditions that the mixture gas contains no external EGR gas and is at a stoichiometric air-fuel ratio. The EGR ratio estimator estimates an EGR ratio based on a pressure difference between an actual in-cylinder pressure detected at the reference crank angle by the in-cylinder pressure sensor, and the calculated reference in-cylinder pressure. The controller controls the internal combustion engine according to the estimated EGR ratio.

According to another aspect of the present invention, a control device for an internal combustion engine, the control device includes a cylinder pressure sensor, a driving condition detector, a reference crank angle setter, a reference cylinder pressure calculator, an EGR ratio estimator, and a controller. The cylinder pressure sensor detects a cylinder pressure inside a cylinder into which fuel is directly injected. The driving condition detector detects a driving condition in the internal combustion engine. The reference crank angle setter calculates, according to the driving condition detected by the driving condition detector, a reference crank angle immediately before which mixture gas starts combusting in the cylinder. The reference cylinder pressure calculator calculates a reference cylinder pressure in the cylinder at the reference crank angle based on temperature characteristics of a heat capacity ratio of the mixture gas under a condition that the mixture gas contains no external EGR gas and that the mixture gas has a stoichiometric air-fuel ratio. The EGR ratio estimator calculates an EGR ratio based on a pressure difference between the reference cylinder pressure and the cylinder pressure detected by the cylinder pressure sensor at the reference crank angle. The controller controls the internal combustion engine according to the EGR ratio.

DESCRIPTION OF THE EMBODIMENT

Hereinafter, a preferable embodiment of the present disclosure is described in detail with reference to the drawings. As illustrated inFIG. 1, an internal combustion engine (hereinafter, referred to as “engine”)3to which the present disclosure is applied is a gasoline engine including, for example, four cylinders3a, and is mounted as a power source on a vehicle (not illustrated).

Each of the cylinders3aof the engine3is provided with a fuel injection valve (hereinafter, referred to as “injector”)4and a spark plug5which face a combustion chamber (not illustrated) of the cylinder3a. The injector4is of a type of directly injecting the fuel into the combustion chamber. Upon electric spark from the spark plug5, the mixture gas of the fuel and the air is ignited and combusted. The fuel injection amount and the fuel injection timing from the injector4and the ignition timing IGLOG of the spark plug5are controlled in accordance with control signals from an electronic control unit (hereinafter, referred to as “ECU”)2(seeFIG. 2).

Here, the “mixture gas” in this embodiment is an in-cylinder gas charged in the cylinder3aand used for combustion, and contains an external exhaust gas recirculation (external EGR) gas in the case where an EGR device14described later performs external EGR.

Each of the cylinders3aof the engine3is provided with an in-cylinder pressure sensor51which detects a pressure inside the cylinder3a(in-cylinder pressure). In this embodiment, the in-cylinder pressure sensor51is a built-in sensor in the injector. Thus, a pressure detection element which faces the combustion chamber and picks up the in-cylinder pressure, an amplifier circuit which amplifies signals from the pressure detection element, and others are assembled and integrated into the injector4, although not illustrated. A detection signal indicating an in-cylinder pressure PCYL detected by the in-cylinder pressure sensor51is inputted to the ECU2.

In addition, the engine3includes a variable intake phase mechanism11, a variable exhaust phase mechanism12, a turbocharger13, the EGR device14, and so forth.

The variable intake phase mechanism11varies a relative phase CAIN of an intake valve (not illustrated) with respect to a crankshaft (not illustrated) of the engine3(hereinafter, such phase is referred to as “intake phase”) steplessly, and includes an intake phase control motor11a(seeFIG. 2) and others. In accordance with a control signal from the ECU2, the intake phase control motor11arotates an intake camshaft (not illustrated) relative to the crankshaft to change a relative angle between the two, and thereby varies the intake phase CAIN steplessly.

Similarly, the variable exhaust phase mechanism12varies a relative phase CAEX of an exhaust valve (not illustrated) with respect to the crankshaft of the engine3(hereinafter, such phase is referred to as “exhaust phase”) steplessly, and includes an exhaust phase control motor12a(seeFIG. 2) and others. In accordance with a control signal from the ECU2, the exhaust phase control motor12arotates an exhaust camshaft (not illustrated) relative to the crankshaft to change a relative angle between the two, and thereby varies the exhaust phase CAEX steplessly.

These variable intake phase mechanism11and variable exhaust phase mechanism12are used to vary the intake phase CAIN and the exhaust phase CAEX, respectively, thereby controlling the valve open-close timings of the intake valve and the exhaust valve, and also controlling internal EGR with valve overlap in which the intake valve and the exhaust valve are both open.

The turbocharger13includes a compressor21provided in an intake passage6and a turbine23provided in an exhaust passage7and integrally coupled to the compressor21via a shaft22. When the turbine23is driven by the exhaust gas flowing through the exhaust passage7, the compressor21rotates integrally with the turbine23and thereby turbo-charges the intake air. Meanwhile, a boost pressure is adjusted by controlling a wastegate valve (not illustrated) and so forth in accordance with a control signal from the ECU2.

In the intake passage6, an intake throttle valve25, the compressor21of the turbocharger13, an inter cooler26which cools the intake air with a temperature increased by turbocharging, and a throttle valve27are provided in this order from the upstream side of the intake air. The intake throttle valve25generates a negative pressure to introduce the external EGR gas to the downstream side of the intake throttle valve25, and the valve lift of the intake throttle valve25is controlled by means of an LP actuator25ain accordance with a control signal from the ECU2.

The throttle valve27is disposed upstream of an intake manifold6aof the intake passage6. The valve lift of the throttle valve27is controlled by means of a TH actuator27ain accordance with a control signal from the ECU2, and thereby an amount of in-cylinder gas to be taken into the cylinder3ais controlled.

A three way catalyst28is provided downstream of the turbine23of the exhaust passage7. The three way catalyst28in an active status purifies the exhaust gas by oxidizing HC and CO and also reducing NOxin the exhaust gas.

The EGR device14recirculates a portion of the exhaust gas, which is discharged from the cylinders3ato the exhaust passage7, as an external EGR gas back to the intake passage6via an EGR passage41. As illustrated inFIG. 1, the EGR passage41is connected to the exhaust passage7at a position downstream of the turbine23and the three way catalyst28and is connected to the intake passage6at a position between the compressor21and the intake throttle valve25. With this structure, the external EGR gas is taken out of the exhaust gas after driving of the turbine23, and accordingly has a relatively low pressure. In other words, the EGR device14is configured as what is termed as a low-pressure EGR device.

In the middle of the EGR passage41, an EGR valve42and an EGR cooler43which cools the external EGR gas are provided. The valve lift of the EGR valve42is controlled by means of an EGR actuator42ain accordance with a control signal from the ECU2, and thereby the amount of the external EGR gas is controlled.

In addition, the engine3is provided with the following various sensors (seeFIG. 2) in addition to the aforementioned in-cylinder pressure sensor51in order to detect driving conditions of the engine3.

Along with rotation of the crankshaft, a crank angle sensor52outputs pulse signals called a CRK signal and a TDC signal to the ECU2at every predetermined respective crank angles. The CRK signal is outputted at every predetermined crank angle (for example, 0.5 degrees). The ECU2calculates the number NE of revolutions of the engine3(hereinafter, referred to as “number of engine revolutions”) based on the CRK signal.

Meanwhile, the TDC signal is a signal indicating that a piston (not illustrated) of the engine3is located at a predetermined crank angle position near an intake top dead center (TDC) in any of the cylinders3a, and is outputted at every 180 degrees of the crank angle in the case where the engine3is a four-cylinder engine as in this embodiment. From the TDC signal and the CRK signal, the ECU2calculates the crank angle CA on the basis of the output timing of the TDC signal for each of the cylinders3a. In addition, from the TDC signal and the CRK signal, The ECU2calculates a crank angle stage FISTG (=0 to 23) at every predetermined crank angle (for example, 30 degrees) and assigns the calculated stage FISTG.

Moreover, the intake camshaft to which the variable intake phase mechanism11is attached, and the exhaust camshaft to which the variable exhaust phase mechanism12is attached are provided with an intake phase sensor53and an exhaust phase sensor54, respectively. Along with rotation of the intake camshaft, the intake phase sensor53outputs a pulse signal called a CAMIN signal to the ECU2at every predetermined cam angle (for example, 0.5 degrees). The ECU2calculates the intake phase CAIN based on the CAMIN signal and the CRK signal. Similarly, along with rotation of the exhaust camshaft, the exhaust phase sensor54outputs a pulse signal called a CAMEX signal to the ECU2at every predetermined cam angle (for example, 0.5 degrees). The ECU2calculates the exhaust phase CAEX based on the CAMEX signal and the CRK signal.

In addition, in the intake passage6, an airflow sensor55is provided upstream of the intake throttle valve25, and an intake pressure sensor56and an intake air temperature sensor57are provided in an intake chamber6blocated downstream of the throttle valve27. The airflow sensor55detects an amount GAIR of air (fresh air) to be taken into the cylinders3a(intake air amount), the intake pressure sensor56detects an absolute pressure of an intake pressure PBA, and the intake air temperature sensor57detects a temperature TA of intake air which contains the external EGR gas and is to be taken into the cylinders3a(intake air temperature). These detection signals are inputted to the ECU2.

An LAF sensor58is provided between the turbine23and the three way catalyst28in the exhaust passage7. The LAF sensor58successively detects an oxide concentration in the exhaust gas to flow into the three way catalyst28in a wide air-fuel ratio range including a stoichiometric ratio, and outputs the detection signal to the ECU2. The ECU2calculates an equivalent ratio KACT of the exhaust gas based on this detection signal.

Further, the ECU2receives a detection signal indicating a temperature TW of cooling water which cools the engine3(hereinafter, referred to as “engine water temperature”) from a water temperature sensor59, and a detection signal indicating a press-down amount AP of an accelerator pedal (not illustrated) of the vehicle (hereinafter, referred to as “accelerator position”) from an accelerator position sensor60.

The ECU2is formed of a microcomputer including a CPU, a RAM, a ROM, an I/O interface, and others (all of which are not illustrated). The ECU2determines the driving conditions of the engine3based on the detection signals and others from the various sensors described above, and performs engine control including control of the fuel injection amount of the injector4, the ignition timing IGLOG of the spark plug5, and the like.

In this embodiment, the ECU2estimates an EGR ratio R_EGR of the mixture gas charged in each of the cylinders3a, and controls the ignition timing IGLOG based on the estimated EGR ratio R_EGR, in particular. It should be noted that the EGR ratio R_EGR of the mixture gas is defined as a ratio of the EGR gas amount to the total amount of the mixture gas (in-cylinder gas).

In this embodiment, the ECU2serves as a reference crank angle setter, a reference in-cylinder pressure calculator, an EGR ratio estimator, a controller, an initial crank angle acquirer, and an initial in-cylinder temperature acquirer.

FIG. 3demonstrates a main flow of estimation processing of an EGR ratio R_EGR executed by the ECU2. This processing is executed for each of the cylinders3aand repeated in the same cycles as cycles of switching the aforementioned crank angle stages FISTG from one to another (for example, at every 30 degrees of the crank angle). Here, processing directly related to the in-cylinder pressure PCYL detected by the in-cylinder pressure sensor51is executed independently of this estimation processing in the same cycles as the generation cycles of the CRK signal (for example, at every 0.5 degrees of the crank angle), and the detected in-cylinder pressure PCYL is stored while being associated with the crank angle CA, for example.

In the estimation processing ofFIG. 3, firstly in step1(denoted by “S1” inFIG. 3; the same applies to the following steps), the ECU2determines whether or not the crank angle stage FISTG is equal to a first predetermined value STG1corresponding to an intake top dead center (TDC). If the determination result is YES and the concerned cylinder3ais in a stage immediately after a shift to an intake stroke, the ECU2acquires intake-related parameters (step2). Specifically, the intake air temperature TA, the engine water temperature TW, and the exhaust phase CAEX are read as the intake-related parameters, and stored into a predetermined area in the RAM of the ECU2. Then, the ECU2terminates this processing.

If the determination result in above step1is NO, the ECU2determines whether or not the crank angle stage FISTG is equal to a second predetermined value STG2corresponding to a compression bottom dead center (BDC) (step3). If the determination result is YES and the concerned cylinder3ais in a stage immediately after a shift to a compression stroke, the ECU2acquires compression-related parameters (step4). Specifically, the intake pressure PBA, the number of engine revolutions NE, and the intake phase CAIN detected, and the ignition timing IGLOG set at this time point are read as the compression-related parameters, and stored into a predetermined area in the RAM of the ECU2.

Next, the ECU2executes setting processing of a reference crank angle CA_REF (step5). This setting processing involves predicting a timing immediately before the start of combustion of the mixture gas, and setting the predicted timing as the reference crank angle CA_REF.FIG. 4demonstrates a sub-routine of this setting processing.

In this setting processing, firstly in step21, the ECU2calculates a retard correction amount ΔC_CA by searching a predetermined map (not illustrated) based on the intake pressure PBA and the number of engine revolutions NE detected in foregoing step4. The retard correction amount ΔC_CA is equivalent to a combustion time lag until the ignited mixture gas starts combustion after the ignition operation by the spark plug5at the ignition timing IGLOG, and is expressed in degrees of the crank angle. The lower the intake pressure PBA, the later the mixture gas starts combustion. Then, the higher the number of engine revolutions NE, the larger the crank angle corresponding to the same combustion time lag. For these reasons, the retard correction amount ΔC_CA in the aforementioned map is set to become larger as the intake pressure PBA becomes lower and as the number of engine revolutions NE becomes higher.

Then, the ECU2sets the reference crank angle CA_REF to a value obtained by subtracting the retard correction amount ΔC_CA from the ignition timing IGLOG acquired in above step4(step22). Here, the reference crank angle CA_REF is expressed in degrees with the origin (0 degrees) set at the compression TDC of each cylinder3awhile an angle on the timing-advance side is expressed in positive degrees (seeFIG. 10).

Next, the ECU2determines whether or not the set reference crank angle CA_REF is smaller than 0 degrees corresponding to the compression TDC (step23). If the determination result is NO, in other words, if the reference crank angle CA_REF is equivalent to the compression TDC or is on the timing-advance side of the compression TDC, the ECU2just terminates this processing.

On the other hand, if the determination result in step23is YES, and the reference crank angle CA_REF is on the timing-retard side of the compression TDC, the ECU2restricts the reference crank angle CA_REF to 0 degrees corresponding to the compression TDC (step24), and terminates this processing.

Returning toFIG. 3, in step6following step5described above, the ECU2executes calculation processing of a reference in-cylinder pressure P_REF. This reference in-cylinder pressure P_REF is an in-cylinder pressure expected to occur at the aforementioned reference crank angle under the conditions that the mixture gas contains no external EGR gas, and is at the stoichiometric air-fuel ratio. The calculation processing thereof will be described later in detail.

Thereafter, the ECU2executes calculation processing of the EGR coefficient C_EGR (step7), and terminates this processing. Since the EGR ratio R_EGR and a pressure difference ΔP (a difference between an actual in-cylinder pressure P_CPS described later and the reference in-cylinder pressure P_REF) are found to have a linear (proportional) relationship therebetween as illustrated in FIG.5, the EGR coefficient C_EGR is defined as a slope (=R_EGR/ΔP) of the line representing the relationship. The calculation processing thereof will be described later in detail.

If the determination result in foregoing step3is NO, the ECU2determines whether or not the crank angle stage FISTG is equal to a third predetermined value STG3corresponding to a compression top dead center (TDC) (step8). If the determination result is NO, the ECU2just terminates the estimation processing. On the other hand, if the determination result is YES and the concerned cylinder3ais in a stage immediately after the end of the compression stroke, the ECU2reads the in-cylinder pressure PCYL detected at the reference crank angle CA_REF set in step5from the RAM, and thereby acquires the read in-cylinder pressure PCYL as an actual in-cylinder pressure P_CPS (step9).

Next, the ECU2calculates a difference between the acquired actual in-cylinder pressure P_CPS and the reference in-cylinder pressure P_REF calculated in step6(=P_CPS−P_REF) (step10). Then, the ECU2multiplies the calculated pressure difference ΔP by the EGR coefficient C_EGR calculated in step7to thereby calculate the EGR ratio R_EGR of the mixture gas (step11), and terminates the estimation processing.

Next, with reference toFIG. 6, description is provided for the calculation processing of the reference in-cylinder pressure P_REF executed in step6inFIG. 3. In this calculation processing, firstly in step31, the ECU2calculates a valve-closing timing IVC of the intake value (hereinafter, referred to as “intake valve-closing timing”) from the intake phase CAIN acquired in foregoing step2. As is the case with the foregoing reference crank angle CA_REF, the intake valve-closing timing IVC is expressed by a crank angle with the origin (0 degrees) set at the compression TDC while an angle on the timing advance side is expressed in positive degrees.

If this intake valve-closing timing IVC is set to a crank angle during the compression stroke, the intake valve-closing timing IVC is equivalent to the crank angle at a compression start time (initial crank angle), because the compression of the mixture gas actually starts at a time when the intake valve. Meanwhile, the intake pressure PBA is equivalent to the in-cylinder pressure at the compression start time (initial in-cylinder pressure).

Next, the ECU2calculates an initial in-cylinder temperature T_STRT by searching a predetermined map (not illustrated) based on the intake air temperature TA, the intake phase CAIN, and the exhaust phase CAEX (step32). Here, the initial in-cylinder temperature T_STRT is a temperature inside the cylinder3aat the compression start time. Among the foregoing parameters, the intake phase CAIN and the exhaust phase CAEX are used to reflect an increase in the in-cylinder temperature depending on an internal EGR amount in the case where internal EGR is performed by way of valve overlap of the intake valve and the exhaust valve. Accordingly, the initial in-cylinder temperature T_STRT in the aforementioned map is set to become a higher value, as the intake air temperature TA becomes higher and as the intake phase CAIN and the exhaust phase CAEX bring about larger valve overlap.

In next step33, the ECU2calculates the reference in-cylinder pressure P_REF by searching a reference in-cylinder pressure map illustrated inFIG. 7based on the reference crank angle CA_REF, the intake valve-closing timing IVC, the initial in-cylinder temperature T_STRT, and the intake pressure PBA. This reference in-cylinder pressure map is explained hereinbelow.

First, the mixture gas (in-cylinder gas) charged in the cylinder3ais explained in terms of the heat capacity ratio and the state change during the compression stroke. The heat capacity ratio κ of the mixture gas is expressed by following Formula (1) by using a specific heat capacity at constant pressure Cp and a gas constant R, and the specific heat capacity at constant pressure Cp is expressed by following Formula (2):

κ=Cp(Cp-R),(1)
where κ denotes the heat capacity ratio of the mixture gas, Cpdenotes the specific heat capacity at constant pressure, and R denotes the gas constant; and

As presented in Formula (2), the heat capacity ratio κ of the mixture gas varies depending on the composition of the mixture gas (components and the numbers of moles of the components). In addition, as presented inFIG. 8, the heat capacity ratio of each component of the mixture gas has a temperature property in which the heat capacity ratio decreases as the temperature increases. Thus, the heat capacity ratio κ of the mixture gas containing these components also has a similar temperature property. Moreover, as presented inFIG. 9, if the mixture gas contains the EGR gas, the composition of the mixture gas is changed, and the CO2component in the EGR gas is added. Hence, the heat capacity ratio κ of the mixture gas has a property of increasing due to the addition of the CO2component.

Meanwhile, the state change of the mixture gas during the compression stroke is regarded as an adiabatic compression change, that is, a polytropic change. Thus, the in-cylinder temperature Taat the crank angle CA=a is expressed by following Formula (3):

Ta=Ta-1⁡(Va-1Va)(κa-1-1),(3)
where Tadenotes an in-cylinder temperature at CA=a, V denotes a cylinder volume, Vadenotes a cylinder volume at CA=a, and κadenotes a heat capacity ratio κ at CA=a.

As expressed by Formula (3), the in-cylinder temperature T is a function of the heat capacity ratio κ, and the heat capacity ratio κ of the mixture gas is a function of the in-cylinder temperature T as described above. Hence, in order to precisely obtain the heat capacity ratio κ and the in-cylinder temperature T, Formulas (1) and (2) and Formula (3) are iteratively calculated by mutually using a calculation result of Formula (3) and calculation results of Formulas (1) and (2), respectively. As a result, the in-cylinder temperature Tθat the crank angle CA=the final crank angle θ (final in-cylinder temperature) is expressed by following Formula (4):

Tθ=T0⁡(V0V1)(κ0-1)⨯(V1V2)(κ1-1)⨯…⨯(Vθ-1Vθ)(κθ-1-1),(4)
where Tθdenotes an in-cylinder temperature at CA=θ (final in-cylinder temperature), T0denotes an initial in-cylinder temperature, V0denotes an initial cylinder volume, Vθdenotes a cylinder volume at CA=θ (final cylinder volume), bκ0denotes an initial heat capacity ratio of the mixture gas, and κθdenotes a heat capacity ratio of the mixture gas at CA=θ.

The in-cylinder pressure Paat the crank angle CA=a is expressed by following Formula (5), and the in-cylinder pressure Pθat the crank angle CA=θ (final in-cylinder pressure) is expressed by following Formula (6) based on Formula (5):

Pa=Pa-1⁡(Va-1Va)κa-1,(5)
where Padenotes the in-cylinder pressure at CA=a; and

As expressed by Formula (6), the final in-cylinder pressure Pθis a function of the initial in-cylinder pressure P0, the initial cylinder volume V0, the final cylinder volume Vθ, and the iteratively-calculated heat capacity ratio κ. Then, the heat capacity ratio κ is a function of the iteratively-calculated in-cylinder temperature T, whereas the in-cylinder temperature T is a function of the initial in-cylinder temperature T0and the heat capacity ratio κ. In addition, since the cylinder volume V is uniquely obtained from the crank angle CA, the initial cylinder volume V0and the final cylinder volume Vθcan be replaced with the initial crank angle CA0and the final crank angle CAθ, respectively.

On the basis of the above, the final in-cylinder pressure Pθis obtained as a function of the initial in-cylinder pressure P0, the initial in-cylinder temperature T0, the initial crank angle CA0and the final crank angle CAθ, provided that the composition of the mixture gas in Formula (2) is given.

The aforementioned reference in-cylinder pressure map is based on the relationships described above, and is configured as presented inFIG. 7, i.e., configured to receive, as input parameters, the intake pressure PBA, the initial in-cylinder temperature T_STRT, and the intake valve-closing timing IVC respectively corresponding to the initial in-cylinder pressure P0, the initial in-cylinder temperature T0, and the initial crank angle CA0, and the reference crank angle CA_REF corresponding to the final crank angle CAθ, and to obtain the reference in-cylinder pressure P_REF corresponding to the final in-cylinder pressure Pθas an output. More specifically, given various conditions of the aforementioned four input parameters, the reference in-cylinder pressure P_REF is calculated based on Formulas (1) to (6) in advance, and the reference in-cylinder pressure map is formed by mapping the calculation results to the input parameters.

As for the composition of the mixture gas, a condition where the mixture gas contains no external EGR gas, a condition of an internal EGR amount, and a condition where the mixture gas is at the stoichiometric air-fuel ratio are given. The first condition is given because an external EGR amount contained in the mixture gas through the external EGR can hardly be known due to a time lag in introduction of the external EGR gas to the cylinder3a. In contrast, unlike the external EGR, the internal EGR introduces the internal EGR gas with almost no time lag, and thus the internal EGR amount is approximately determined by the foregoing initial conditions including the intake valve-closing timing IVC, and therefore is given as the condition.

To be more specific, the internal EGR amount is calculated by way of simulation or the like based on the intake pressure PBA, the initial in-cylinder temperature T_STRT, and the intake valve-closing timing IVC. Then, in foregoing Formula (2), the number of moles nCO2of the CO2component and the number of moles nH2Oof the H2O component of the exhaust gas components are set depending on the calculated internal EGR amount, and the numbers of moles nXof the other components are set at a ratio corresponding to the stoichiometric air-fuel ratio. Given the foregoing conditions of the composition of the mixture gas, the reference in-cylinder pressure P_REF is calculated in advance based on Formulas (1) to (6) under various conditions of the foregoing four input parameters, and the reference in-cylinder pressure map is formed by mapping the calculation results to the input parameters.

FIGS. 10 to 12present setting examples of the reference in-cylinder pressure P_REF relative to the input parameters in the reference in-cylinder pressure map. As presented inFIG. 10, the reference in-cylinder pressure P_REF is set to become a larger value, as the value of the reference crank angle CA_REF becomes closer to 0, that is, as the reference crank angle CA_REF becomes closer to the compression TDC. In addition, the reference in-cylinder pressure P_REF is set to become a larger value, as the value of the intake valve-closing timing IVC becomes larger, that is, as the valve opening timing of the intake valve in the compression stroke becomes earlier. This is because the reference crank angle CA_REF closer to the compression TDC and the earlier valve-closing timing of the intake valve bring a longer actual compression period of the mixture gas, and result in a higher final in-cylinder pressure.

Moreover, as presented inFIG. 11, the reference in-cylinder pressure P_REF is set to become a smaller value as the initial in-cylinder temperature T_STRT becomes higher. This is because the higher initial in-cylinder temperature T_STRT makes the in-cylinder temperature higher and accordingly the heat capacity ratio κ of the mixture gas lower, which in turn lowers the increase rate of the in-cylinder pressure.

In addition, as presented inFIG. 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 respectively correspond to the final in-cylinder pressure Pθand the initial in-cylinder pressure P0, which have a proportional relationship therebetween (see Formula (6)).

As described above, in step33inFIG. 6, the ECU2calculates the reference in-cylinder pressure P_REF by searching the aforementioned reference in-cylinder pressure map based on the foregoing four parameters. In next step34, the ECU2calculates a heat-transfer correction coefficient K_HT by searching a predetermined map based on the number of engine revolutions NE and the engine water temperature TW. This heat-transfer correction coefficient K_HT is intended to compensate for influence of heat transferred between the inside and the outside of the cylinder3a.

Then, the ECU2calculates the final reference in-cylinder pressure P_REF by multiplying the reference in-cylinder pressure P_REF calculated in step33by the heat-transfer correction coefficient K_HT (step35), and terminates this processing.

Hereinbelow, with reference toFIG. 13, description is provided for the calculation processing of the EGR coefficient C_EGR executed in step7ofFIG. 3. As described above, the EGR coefficient C_EGR is defined as the slope of the EGR ratio R_EGR relative to the pressure difference ΔP between the actual in-cylinder pressure P_CPS and the reference in-cylinder pressure P_REF (seeFIG. 5), and is used to calculate the EGR ratio R_EGR. Since the aforementioned slope is found to have a property of varying depending on the intake conditions and the compression conditions, the EGR coefficient C_EGR is calculated in this calculation processing.

In this calculation processing, firstly in step41, the ECU2acquires the reference crank angle CA_REF, the intake valve-closing timing IVC, the initial in-cylinder temperature T_STRT, and the intake pressure PBA. These four parameters represent the aforementioned intake conditions and compression conditions, and are the same as the four input parameters for the aforementioned reference in-cylinder pressure map. Thus, the acquisition of the parameters in step41may be done by reading the data acquired in the calculation processing of the reference in-cylinder pressure P_REF in FIG.6.

Then, the ECU2calculates the EGR coefficient C_EGR by searching an EGR coefficient map illustrated inFIG. 14based on the acquired four parameters (step42), and terminates this calculation processing. This EGR coefficient map is formed by calculating the EGR coefficient C_EGR in advance based on Formulas (1) to (6) under various conditions of the foregoing four input parameters, and mapping the calculation results to the input parameters.

FIGS. 15 to 17present setting examples of the EGR coefficient C_EGR relative to the input parameters in the EGR coefficient map. As presented inFIG. 15, the EGR coefficient C_EGR is set to become a smaller value, as the reference crank angle CA_REF becomes closer to the compression TDC and as the valve-closing timing of the intake valve in the compression stroke becomes earlier. This is because the reference crank angle CA_REF closer to the compression TDC and the earlier valve-closing timing of the intake valve bring a longer actual compression period of the mixture gas, which in turn results in a larger pressure difference ΔP, so that the EGR coefficient C_EGR accordingly becomes smaller.

Moreover, as presented inFIG. 16, the EGR coefficient C_EGR is set to become a smaller value as the initial in-cylinder temperature T_STRT becomes higher, for the following reason. Specifically, among the components of the mixture gas, the fuel has relatively great temperature-dependent variation in the specific heat capacity at constant pressure Cp, and therefore makes relatively great contribution to the temperature property of the heat capacity ratio κ of the mixture gas. Meanwhile, in the case where the EGR ratio R_EGR increases, the ratio of the fuel decreases accordingly, which then lowers the contribution of the fuel and consequently makes the temperature-dependent variation in the heat capacity ratio κ smaller. Thus, as the initial in-cylinder temperature T_STRT becomes higher, the variation in the heat capacity ratio κ during the compression becomes more greatly to bring a larger pressure difference ΔP, so that the EGR coefficient C_EGR accordingly becomes smaller.

Further, as presented inFIG. 17, the EGR coefficient C_EGR is set to become a smaller value as the intake pressure PBA becomes higher. This is because, as the intake pressure PBA being the initial in-cylinder pressure becomes higher, the actual in-cylinder pressure P_CPS and the pressure difference ΔP increase proportionally, and the EGR coefficient C_EGR accordingly becomes smaller.

Hereinafter, with reference toFIG. 18, description is provided for ignition timing control processing using the EGR ratio R_EGR. This control processing is executed for each of the cylinder3ain synchronization with the generation of the TDC signal. In this control processing, firstly in step51, the ECU2calculates a base ignition timing IG_BASE by searching a predetermined map (not illustrated) based on the number of engine revolutions NE and a demand torque TRQCMD. The demand torque TRQCMD herein is calculated based on the accelerator position AP and the number of engine revolutions NE.

Subsequently, the ECU2calculates an EGR correction amount ΔIGEGR by searching a predetermined map (not illustrated) based on the estimated EGR ratio R_EGR (step52).

Then, the ECU2calculates a correction amount ΔIGTTL based on the engine water temperature TW and the number of engine revolutions NE (step53). The correction amount ΔIGTTL is intended for correction due to factors other than the EGR ratio R_EGR.

Lastly, the ECU2calculates the ignition timing IGLOG by adding the EGR correction amount ΔIGEGR and the correction amount ΔIGTTL to the base ignition timing IG_BASE (step54), and terminates the control processing.

As described above, according to this embodiment, the reference in-cylinder pressure P_REF expected to occur at the reference crank angle CA_REF is calculated based on the temperature property of the heat capacity ratio κ of the mixture gas under the conditions where the mixture gas contains no external EGR gas and is at the stoichiometric air-fuel ratio. Then, the EGR ratio of the mixture gas is calculated based on the pressure difference ΔP between the reference in-cylinder pressure P_REF and the actual in-cylinder pressure P_CPS detected at the reference crank angle CA_REF. This makes it possible to estimate the EGR ratio R_EGR while reflecting the temperature property of the heat capacity ratio κ of the mixture gas.

In addition, the reference crank angle CA_REF is the crank angle immediately before the start of combustion of the mixture gas. Such setting of the reference crank angle CA_REF may enable the actual in-cylinder pressure P_CPS to be acquired in the state where the mixture gas is yet to burn and keeps the polytropic state change, and also may ensure a large pressure difference ΔP between the actual in-cylinder pressure P_CPS and the reference in-cylinder pressure P_REF. Thus, the temperature property of the heat capacity ratio κ of the mixture gas can be advantageously reflected based on that pressure difference ΔP, so that the EGR ratio R_EGR may be estimated with high accuracy. Then, the ignition timing IGLOG may be controlled appropriately using the EGR ratio R_EGR thus estimated with high accuracy.

The actual in-cylinder pressure P_CPS being an actual pressure inside the cylinder3aand the reference in-cylinder pressure P_REF thereof are used as the parameters to estimate the EGR ratio R_EGR. Thus, even if the EGR device14is the low-pressure EGR device, highly accurate estimation of the EGR ratio R_EGR may be achieved without receiving an influence of a time lag of the external EGR gas.

Moreover, the reference crank angle CA_REF is set by using the ignition timing IGLOG, the intake pressure PBA, and the number of engine revolutions NE, and thus can be appropriately set depending on the actual driving conditions of the engine3. This enables the reference in-cylinder pressure P_REF and the actual in-cylinder pressure P_CPS at the reference crank angle CA_REF to be appropriately obtained.

In addition, if the set reference crank angle CA_REF is on the timing-retard side of 0 degrees corresponding to the compression TDC, the reference crank angle CA_REF is restricted to 0 degrees. This restriction may keep the actual in-cylinder pressure P_CPS from decreasing due to the influence of knocking and the like after the compression TDC, and therefore a pressure difference ΔP obtained between the actual in-cylinder pressure P_CPS and the reference in-cylinder pressure P_REF may be made so large that the estimation accuracy of the EGR ratio R_EGR may be kept high.

Further, the reference in-cylinder pressure P_REF can be calculated appropriately based on the reference crank angle CA_REF, the intake valve-closing timing IVC equivalent to the initial crank angle at the compression start time, the initial in-cylinder temperature T_STRT, and the intake pressure PBA equivalent to the initial in-cylinder pressure. Then, the reference in-cylinder pressure P_REF thus calculated is corrected depending on the number of engine revolutions NE and the engine water temperature TW to thus compensate for the influence of heat transferred between the inside and the outside of the cylinder3a.

Still further, appropriate calculation of the EGR coefficient C_EGR can be achieved by reflecting the intake and compression conditions of the mixture gas based on the same four parameters (the reference crank angle CA_REF, the intake valve-closing timing IVC, the initial in-cylinder temperature T_STRT, and the intake pressure PBA) as those used to calculate the reference in-cylinder pressure P_REF. Then, the EGR coefficient C_EGR thus calculated is multiplied by the pressure difference ΔP, so that the EGR ratio R_EGR can also be estimated appropriately.

Furthermore, the in-cylinder pressure sensor51is formed of the pressure detection element and the amplifier circuit which are integrated into the injector4, and thus is less susceptible to noise due to the ignition operation and noise due to injection operations by the injectors4of the other cylinders3a. This enables the in-cylinder pressure sensor51to achieve higher detection accuracy of the actual in-cylinder pressure P_CPS, and thereby leads to further improvement of estimation accuracy of the EGR ratio R_EGR.

Hereinafter, with reference toFIG. 19, description is provided for a modification of the calculation processing of the reference in-cylinder pressure P_REF. In this modification, since the reference in-cylinder pressure P_REF has the aforementioned proportional relationship with the intake pressure PBA (FIG. 12), the intake pressure PBA is excluded from the input parameters for the reference in-cylinder pressure map, and the map value obtained from the reference in-cylinder pressure map is corrected by using the intake pressure PBA. This calculation processing is executed instead of the processing inFIG. 6. InFIG. 19, steps involving the same processing contents as those inFIG. 6are indicated by the same step numbers.

In this calculation processing, the ECU2firstly executes steps31and32which are the same as those inFIG. 6to calculate the intake valve-closing timing IVC and the initial in-cylinder temperature T_STRT. Then, the ECU2calculates the reference in-cylinder pressure P_REF by searching the reference in-cylinder pressure map based on the reference crank angle CA_REF, the intake valve-closing timing IVC, and the initial in-cylinder temperature T_STRT (step301). Note that, in this reference in-cylinder pressure map, the initial in-cylinder pressure at the compression start time is treated as a constant, which is set to a reference atmospheric pressure PATM (760 mmHg).

Then, the ECU2sets an intake pressure correction coefficient K_PB to a quotient of the intake pressure PBA divided by the reference atmospheric pressure PATM (step302), and multiplies the reference in-cylinder pressure P_REF calculated in step301by the intake pressure correction coefficient K_PB to calculate the corrected reference in-cylinder pressure P_REF (step303).

The following processing contents are the same as those inFIG. 6. Thus, the ECU2calculates the final reference in-cylinder pressure P_REF by multiplying the reference in-cylinder pressure P_REF calculated in step303by the heat-transfer correction coefficient K_HT calculated based on the number of engine revolutions NE and the engine water temperature TW (steps34,35), and terminates this calculation processing.

The modification described above may achieve calculation of the reference in-cylinder pressure P_REF comparable with that in the case of the calculation processing inFIG. 6, and also may make creation of the reference in-cylinder pressure map easier with decreased less number of input parameters, thereby reducing the workload for the map creation.

It is to be noted that the present disclosure should not be limited to the foregoing embodiment, but may be carried out in other various modes. For example, the above embodiment uses the ignition timing IGLOG, the intake pressure PBA, and the number of engine revolutions NE as the parameters used to calculate the reference crank angle CA_REF, but any other appropriate parameter may be used in addition.

Moreover, in the foregoing embodiment, the initial in-cylinder temperature T_STRT to be used to calculate the reference in-cylinder pressure P_REF and the EGR coefficient C_EGR is calculated based on the intake air temperature TA, the intake phase CAIN, and the exhaust phase CAEX. However, if the internal EGR by valve overlap of the intake and exhaust values is not performed, the intake air temperature TA may be used directly as the initial in-cylinder temperature. In addition, as the initial in-cylinder pressure, the intake pressure PBA is used, but the in-cylinder pressure PCYL detected at the compression start time by the in-cylinder pressure sensor51may be used instead.

The reference in-cylinder pressure P_REF is corrected based on the number of engine revolutions NE and the engine water temperature TW, but may be corrected by additionally using any other appropriate parameter considered to have influence on the heat transfer between the inside and the outside of the cylinder3a.

Further, in the foregoing embodiment, the ignition timing control is performed based on the estimated EGR ratio R_EGR. Instead of or in addition to the ignition timing control, any other kinds of engine control based on the estimated EGR ratio R_EGR may be executed such for example as EGR control via the EGR valve42, intake air amount control via the throttle valve27, and fuel injection control via the injector4.

Furthermore, the EGR device14is the low-pressure EGR device in the foregoing embodiment. Instead of or in addition to the low-pressure EGR device, a high-pressure EGR device may be used. Also in this case, the effects described above may be obtained as well. Then, the in-cylinder pressure sensor51is the built-in component in the injector4, but may be a separate component arranged apart from the injector4, as a matter of course.

Still further, the engine3is the engine for automobile in the foregoing embodiment. However, the present disclosure may be applied to other engines for different purposes, such as an outboard engine in which the crankshaft is arranged in a vertical direction, for example. Moreover, the present disclosure may be altered in details as needed without departing from the spirit of the present disclosure.

A first aspect of the present disclosure provides a control device for an internal combustion engine including an EGR device14that, concurrently with direct injection of a fuel into a cylinder3a, recirculates a portion of an exhaust gas discharged to an exhaust passage7from the cylinder3aback to an intake passage6as an external EGR gas, the controller device including: an in-cylinder pressure sensor51that detects a pressure inside the cylinder3aas an in-cylinder pressure PCYL; a driving condition detector (an intake pressure sensor56, a crank angle sensor52, and an ECU2) that detects driving conditions (an ignition timing IGLOG, an intake pressure PBA, and the number of engine revolutions NE) of an internal combustion engine3; a reference crank angle setter (the ECU2, step5inFIG. 3, andFIG. 4) that obtains a crank angle immediately before start of combustion of a mixture gas charged in the cylinder3adepending on the detected driving conditions of the internal combustion engine3, and sets the obtained crank angle as a reference crank angle CA_REF; a reference in-cylinder pressure calculator (the ECU2, step6inFIG. 3, andFIG. 6) that calculates, as a reference in-cylinder pressure P_REF, a pressure inside the cylinder3aexpected to occur at the set reference crank angle CA_REF, based on a temperature property of a heat capacity ratio of the mixture gas under conditions that the mixture gas contains no external EGR gas and is at a stoichiometric air-fuel ratio; an EGR ratio estimator (the ECU2and steps9to11inFIG. 3) that estimates an EGR ratio R_EGR based on a pressure difference ΔP between an actual in-cylinder pressure detected at the reference crank angle CA_REF by the in-cylinder pressure sensor51, and the calculated reference in-cylinder pressure P_REF; and a controller (the ECU2andFIG. 18) that controls the internal combustion engine3according to the estimated EGR ratio R_EGR.

In this internal combustion engine, the fuel is directly injected into the cylinder, and concurrently a portion of the exhaust gas discharged from the cylinder to the exhaust passage is recirculated as the external EGR gas into the intake passage. In the control device for an internal combustion engine according to the present disclosure, the in-cylinder pressure sensor detects the in-cylinder pressure (the pressure inside the cylinder). Then, the crank angle immediately before the start of combustion of the mixture gas obtained depending on the detected driving conditions of the internal combustion engine is set as the reference crank angle, and the pressure inside the cylinder expected to occur at this reference crank angle is calculated as the reference in-cylinder pressure. This calculation of the reference in-cylinder pressure is done based on the temperature property of the heat capacity ratio of the mixture gas under the conditions that a composition of the mixture gas contains no external EGR gas and is at the stoichiometric air-fuel ratio.

As described above, the heat capacity ratio of the mixture gas is basically determined by the composition of the mixture gas, and has the temperature property in which the heat capacity ratio varies depending on the temperature of the mixture gas. Meanwhile, in the case where the external EGR is performed, there is a time lag until the external EGR gas reaches the inside of the cylinder due to a long passage leading to the cylinder, and therefore the external EGR amount can hardly be known. Against this background, the calculation of the reference in-cylinder pressure based on the temperature property of the heat capacity ratio under the foregoing conditions of the mixture gas enables the reference in-cylinder pressure to be uniquely and appropriately obtained with the temperature property of the heat capacity ratio reflected therein.

In addition, according to the present disclosure, the in-cylinder pressure detected at the reference crank angle is obtained as the actual in-cylinder pressure, and the EGR ratio is estimated based on the pressure difference between the actual in-cylinder pressure and the reference in-cylinder pressure. This actual in-cylinder pressure reflects the actual composition of the mixture gas containing the external EGR gas, and the temperature and the heat capacity ratio depending on the actual composition of the mixture gas. Hence, it is possible to estimate the EGR ratio based on the pressure difference between the actual in-cylinder pressure and the reference in-cylinder pressure.

Moreover, the reference crank angle is the crank angle immediately before the start of combustion of the mixture gas, and is set depending on the detected driving conditions of the internal combustion engine. Such setting of the reference crank angle may enable the actual in-cylinder pressure to be acquired in the state where the mixture gas is yet to burn and keeps the polytropic state change, and also may ensure a large pressure difference between the actual in-cylinder pressure and the reference in-cylinder pressure. Thus, the EGR ratio may be estimated with high accuracy while advantageously reflecting the temperature property of the heat capacity ratio based on this pressure difference. Then, the control of the internal combustion engine according to the EGR ratio thus accurately estimated may achieve the appropriate control of the internal combustion engine.

A second aspect of the present disclosure, which is based on the control device for an internal combustion engine according to the first aspect, may be preferably characterized in that the EGR device14recirculates the external ERG gas from a downstream side of a turbine23of a supercharger (a turbocharger13) in the exhaust passage7back to an upstream side of a compressor21of the supercharger in the intake passage6.

The EGR device configured as described above is what is termed as a low-pressure EGR device, and has such a tendency that the external EGR gas reaches the cylinder with a long time lag due to a relatively long passage through which the external EGR gas needs to flow. In the present disclosure, the EGR ratio is estimated by using, as parameters, the actual in-cylinder pressure that is an actual pressure inside the cylinder, and its reference in-cylinder pressure as described above. Thus, even in the case of the low-pressure EGR device, the EGR ratio may be estimated with high accuracy without being influenced by the time lag of the external EGR gas, which may produce the effect of the present disclosure effectively in particular.

A third aspect of the present disclosure, which is based on the control device for an internal combustion engine according to the first or second aspect, may be preferably characterized in that the driving condition detector detects an ignition timing IGLOG, a pressure of intake air to be taken in the cylinder3a(intake pressure PBA), and the number of revolutions NE of the internal combustion engine3as the driving conditions of the internal combustion engine3, and the reference crank angle setter sets the reference crank angle CA_REF depending on the ignition timing IGLOG, the pressure of intake air, and the number of revolutions NE of the internal combustion engine3thus detected (FIG. 4).

As described above, the reference crank angle is set to the crank angle immediately before the start of combustion of the mixture gas. Meanwhile, the combustion start timing of the mixture gas is directly influenced by the ignition timing, varies depending on the pressure of intake air, and also varies, if expressed by using a crank angle, depending on the number of revolutions of the internal combustion engine. With this configuration, the reference crank angle is set depending on these three detected parameters. This may enable appropriate setting of the reference crank angle depending on the actual driving conditions of the internal combustion engine with the result that the reference in-cylinder pressure and the actual in-cylinder pressure at the reference crank angle may be appropriately obtained. Here, the “detection” of various parameters in the present application includes not only directly detecting a parameter by means of a sensor or the like, but also calculating or estimating the parameter through an arithmetic operation, for example.

A fourth aspect of the present disclosure, which is based on the control device for an internal combustion engine according to the third aspect, may be preferably characterized in that, if the set reference crank angle CA_REF is on a timing-retard side of a compression TDC, the reference crank angle setter restricts the reference crank angle CA_REF to a crank angle corresponding to the compression TDC (0 degrees) (steps23and24inFIG. 4).

For example, as illustrated inFIG. 20, when the ignition timing is on the timing-advance side of the compression TDC and the combustion start timing of the mixture gas is on the timing-retard side of the compression TDC, the actual in-cylinder pressure may decrease (a section indicated by an arrow A inFIG. 20) between the compression TDC and the combustion start mainly due to an influence of knocking. In this case, if the reference crank angle is set on the timing-retard side of the compression TDC, the actual in-cylinder pressure detected at the reference crank angle may decrease, and a large pressure difference may not be obtained between the actual in-cylinder pressure and the reference in-cylinder pressure. As a result, the estimation accuracy of the EGR ratio based on the pressure difference may be lowered.

With the above configuration, if the reference crank angle set as described above is on the timing-retard side of the compression TDC, the reference crank angle is restricted to the crank angle corresponding to the compression TDC. This restriction may keep the actual in-cylinder pressure from decreasing due to the influence of knocking and the like after the compression TDC, and therefore such a large pressure difference may be obtained that the estimation accuracy of the EGR ratio may be kept high.

A fifth aspect of the present disclosure, which is based on the control device for an internal combustion engine according to any one of the first to fourth aspects, may further include an initial crank angle acquirer (an intake phase sensor53and the ECU2) that acquires, as an initial crank angle (intake valve-closing timing IVC), the crank angle at a compression start time when the mixture gas starts to be compressed in a compression stroke; an initial in-cylinder temperature acquirer (an intake air temperature sensor57, an intake phase sensor53, an exhaust phase sensor54, the ECU2, and step32inFIG. 6) that acquires, as an initial in-cylinder temperature T_STRT, a temperature inside the cylinder3aat the compression start time; and an initial in-cylinder pressure acquirer (an intake pressure sensor56) that acquires, as initial in-cylinder pressure (intake pressure PBA), the pressure inside the cylinder3aat the compression start time. The reference in-cylinder pressure calculator may calculate the reference in-cylinder pressure P_REF based on the temperature property of the heat capacity ratio of the mixture gas, depending on the reference crank angle CA_REF, and also the initial crank angle, the initial in-cylinder temperature T_STRT, and the initial in-cylinder pressure thus acquired (step33inFIG. 6).

The reference in-cylinder pressure is the in-cylinder pressure expected to occur at the reference crank angle corresponding to a time point immediately before the start of combustion of the mixture gas. Thus, the reference in-cylinder pressure varies depending on the reference crank angle and also varies depending on the compression start timing of the mixture gas, and the temperature and the pressure of the mixture gas at the compression start time. With this configuration, the reference in-cylinder pressure is calculated based on the reference crank angle, as well as the initial crank angle, the initial in-cylinder temperature, and the initial in-cylinder pressure at the compression start time. Thus, the reference in-cylinder pressure may be calculated appropriately.

A sixth aspect of the present disclosure, which is based on the control device for an internal combustion engine according to the fifth aspect, may further include a number of revolutions detector (a crank angle sensor52) that detects the number of revolutions NE of the internal combustion engine, and a cooling water temperature detector (a water temperature sensor59) that detects a temperature TW of cooling water cooling the internal combustion engine. The reference in-cylinder pressure calculator may correct the reference in-cylinder pressure P_REF depending on the number of revolutions NE of the internal combustion engine3and the temperature TW of the cooling water thus detected (steps34,35inFIG. 6).

With this configuration, the correction of the reference in-cylinder pressure depending on the detected number of revolutions of the internal combustion engine and the detected temperature of the cooling water may compensate for influence of heat transferred between the inside and the outside of the engine.

A seventh aspect of the present disclosure, which is based on the control device for an internal combustion engine according to the fifth or sixth aspect, may preferably be characterized in that the EGR ratio estimator calculates an EGR coefficient C_EGR indicative of a slope of the EGR ratio R_EGR relative to the pressure difference ΔP, based on the temperature property of the heat capacity ratio of the mixture gas, depending on the reference crank angle CA_REF, the initial crank angle, the initial in-cylinder temperature T_STRT, and the initial in-cylinder pressure, and calculates the EGR ratio R_EGR by multiplying the pressure difference ΔP by the calculated EGR coefficient C_EGR (steps7,11inFIG. 3andFIG. 13).

The pressure difference between the actual in-cylinder pressure and the reference in-cylinder pressure and the EGR ratio have a proportional relationship therebetween, and are found to have a property in which a proportional constant (slope) thereof varies depending on the intake conditions and the compression conditions. With this configuration, in estimation of the EGR ratio, the EGR coefficient indicative of the slope of the EGR ratio relative to the pressure difference is firstly calculated depending on the reference crank angle, the initial crank angle, the initial in-cylinder temperature, and the initial in-cylinder pressure. In this way, it is possible to appropriately calculate the EGR coefficient while reflecting the intake and compression conditions of the mixture gas by using the same parameters as those used for calculation of the reference in-cylinder pressure. Then, the pressure difference is multiplied by EGR coefficient thus calculated, so that the EGR ratio may be estimated with high accuracy.

A eight aspect of the present disclosure, which is based on the control device for an internal combustion engine according to any one of the first to seventh aspects, may be preferably characterized in that the in-cylinder pressure sensor51includes a pressure detection element that detects the in-cylinder pressure, and an amplifier circuit that amplifies and outputs a signal outputted from the pressure detection element, and the pressure detection element and the amplifier circuit are integrated into a fuel injection value4that injects the fuel into the cylinder3a.

The in-cylinder pressure sensor configured as described above is formed of the pressure detection element and the amplifier circuit integrated into the fuel injection valve, and thus may be less susceptible to noise due to an ignition operation and noise due to injection operations by the fuel injection values of the other cylinders. This enables the in-cylinder pressure sensor to achieve higher detection accuracy of the actual in-cylinder pressure, and thereby leads to further improvement of estimation accuracy of the EGR ratio.