JP4519164B2 - Internal combustion engine pressure / temperature calculation device - Google Patents

Description

  The present invention relates to a pressure / temperature calculation device for an internal combustion engine.

  The pressure in the intake pipe based on the law of conservation of mass and the law of conservation of energy established between the air flowing into the intake pipe of the internal combustion engine from the atmosphere and the gas (for example, exhaust gas) flowing into the intake pipe other than this air A method for calculating the value is disclosed in Patent Document 1.

JP 2002-147279 A JP 2001-41095 A

  As can be seen from Patent Document 1 as a whole, the method disclosed herein is considered to be a method devised on the assumption that gas flowing into the intake pipe other than air diffuses throughout the intake pipe. However, gas that has flowed into the intake pipe other than air does not diffuse throughout the intake pipe, but has a certain distribution. Therefore, the method disclosed in Patent Document 1 has room for improvement to be made in order to calculate the pressure in the intake pipe more accurately.

  The present invention is to accurately calculate the pressure in the intake pipe of an internal combustion engine.

In order to solve the above problem, according to the first invention, the exhaust passage is connected to the intake passage downstream of the throttle valve via the exhaust gas recirculation passage, and the flow rate of the exhaust gas flowing in the exhaust gas recirculation passage is controlled. In an internal combustion engine in which an exhaust gas flow control valve to be controlled is disposed in an exhaust gas recirculation passage, an air amount calculating means for repeatedly calculating the amount of air passing through a throttle valve, and exhaust gas passing through the exhaust gas flow control valve Exhaust gas amount calculating means for repeatedly calculating the amount of gas, in-cylinder gas amount calculating means for repeatedly calculating the amount of gas flowing into the cylinder, at least the air amount, the exhaust gas amount, and the in-cylinder gas amount And a means for repeatedly calculating the pressure in the intake passage downstream of the throttle valve based on the exhaust gas amount calculating means based on at least the previous pressure in the intake passage. A pressure calculating device for calculating the quantity of exhaust gas, the amount of exhaust gas passing through the exhaust gas flow control valve, the pressure of the same throttle valve in the downstream of the intake passage, the and the cylinder at the time of steady operation The flow rate of air flowing into the cylinder when the flow rate of exhaust gas flowing into the cylinder is zero, and the flow rate of air flowing into the cylinder during steady operation and when the flow rate of exhaust gas flowing into the cylinder is not zero A pressure calculation device that is calculated from the difference between the two is provided.

According to a second aspect, in the first aspect, the difference is two linear function expressions of the pressure in the intake passage downstream of the throttle valve, and the gradients are different from each other and are continuous at the connection point. It is calculated by one linear function formula.

In order to solve the above problem, according to a third invention, in the first invention, the temperature of the exhaust gas that passes through the exhaust gas flow control valve and flows into the intake passage downstream of the throttle valve represents the engine operating state. Calculated from a map with parameters as variables .

  Since it is not necessary to use a large number of sensors, the number of sensors can be reduced, and thus the manufacturing cost can be reduced. Further, since it is not necessary to use a large number of maps, it is possible to reduce the labor for creating the maps.

  Embodiments of the present invention will be described below with reference to the drawings.

  FIG. 1 is an overall view of an internal combustion engine to which the present invention is applied. In the following description, a cylinder injection type spark ignition type internal combustion engine will be described as an example. However, the present invention can be applied to other spark ignition type internal combustion engines and compression self-ignition type internal combustion engines.

  In FIG. 1, 1 is an engine body, 2 is a cylinder block, 3 is a piston, 4 is a cylinder head, 5 is a cylinder (combustion chamber), 6 is an intake valve, 7 is an intake port, 8 is an exhaust valve, and 9 is an exhaust port. Reference numeral 10 denotes a spark plug, 11 denotes a fuel injection valve, and 12 denotes a cavity.

  The intake port 7 is connected to the surge tank 14 via the intake branch pipe 13 for each cylinder 5. The surge tank 14 is connected to an air cleaner 16 via an upstream side intake pipe 15. A throttle valve 18 driven by a step motor 17 is disposed in the upstream side intake pipe 15.

  On the other hand, the exhaust port 9 is connected to an exhaust pipe 19. The exhaust pipe 19 is connected to the exhaust purification device 20.

  The exhaust pipe 19 is connected to the intake branch pipe 13 via an exhaust gas recirculation passage (hereinafter referred to as an EGR passage). Further, an EGR control valve 22 that controls the flow rate of exhaust gas flowing through the EGR passage 21 is disposed in the EGR passage 21. Exhaust gas discharged from the cylinder 5 of the internal combustion engine is again introduced into the cylinder through the EGR passage 21. In the present specification, the exhaust gas recirculated into the cylinder through the EGR passage 21 is also referred to as EGR gas.

  The electronic control unit (ECU) 31 comprises a digital computer, and is connected to each other via a bidirectional bus 32, a RAM (random access memory) 33, a ROM (read only memory) 34, a CPU (microprocessor) 35, an input. A port 36 and an output port 37 are provided.

  An intake pipe pressure sensor 40 for detecting the pressure in the intake branch pipe 13 (hereinafter referred to as intake pipe pressure) is attached to the intake branch pipe 13. The intake pipe pressure sensor 40 generates an output voltage proportional to the intake pipe pressure, and this output voltage is input to the input port 36 via the corresponding AD converter 38.

  The internal combustion engine also has a throttle opening sensor 43 for detecting the opening of the throttle valve 18, the atmospheric pressure around the internal combustion engine, or the pressure of the air taken into the upstream intake pipe 15 (intake pressure). And an atmospheric temperature sensor 45 for detecting the temperature of the atmosphere around the internal combustion engine or the temperature of the air taken into the upstream side intake pipe 15 (intake air temperature). . These sensors 44 and 45 generate output voltages proportional to the atmospheric pressure and the atmospheric temperature, respectively, and these output voltages are input to the input port 36 via the corresponding AD converters 38.

  A load sensor 47 that generates an output voltage proportional to the amount of depression of the accelerator pedal 46 is connected to the accelerator pedal 46. The output voltage of the load sensor 47 is input to the input port 36 via the corresponding AD converter 38.

  The internal combustion engine includes a crank angle sensor 48. For example, the crank angle sensor 48 generates an output pulse every time the crankshaft rotates 30 degrees, and the output pulse is input to the input port 36. The CPU 35 calculates the engine speed from the output pulse of the crank angle sensor 45.

  On the other hand, the output port 37 is connected to the spark plug 10, the fuel injection valve 11, the step motor 17, and the EGR control valve 22 via a corresponding drive circuit 39. The opening degree of the EGR control valve 22 is calculated in the ECU 31 based on a step signal sent from the output port 37 to the EGR control valve 22.

  By the way, the amount of fuel to be injected into the cylinder 5 from the fuel injection valve 10 (hereinafter simply referred to as the fuel injection amount) is based on the amount of air charged in the cylinder 5 and the mixed gas in the cylinder 5. The air-fuel ratio is determined to be the target air-fuel ratio. Therefore, in order to accurately set the air-fuel ratio of the mixed gas in the cylinder 5 to the target air-fuel ratio, it is necessary to accurately grasp the amount of air charged in the cylinder 5 (hereinafter referred to as in-cylinder charged air amount). There is.

  Here, as a method of estimating the cylinder air charge amount, an air flow meter for detecting the mass flow rate of air passing through the throttle valve 18 and other sensors are attached to the internal combustion engine, and the output values of these sensors are used as variables. There is a method of preparing a map for calculating the in-cylinder charged air amount in advance and estimating the in-cylinder charged air amount from the output value of the sensor and the map.

  However, when estimating the in-cylinder charged air amount using a map, in order to accurately estimate the in-cylinder charged air amount, the number of maps and their arguments must be increased. However, if the number of maps is increased, the storage capacity of the ROM 34 for storing these maps must be increased, which increases the cost of the internal combustion engine.

  Furthermore, when estimating the cylinder charge air amount using a map, it is necessary to create a map for each internal combustion engine model, or for each internal combustion engine even if it is the same model internal combustion engine, The map creation effort increases.

  On the other hand, if the number of map arguments is increased in order to more accurately estimate the amount of air charged in the cylinder, the map creation labor is greatly increased.

  Therefore, a method for calculating the in-cylinder charged air amount by numerical calculation using an expression derived from a model instead of a map has been studied, and such a method has already been filed by the present applicant (Japanese Patent Application No. 2001-316350). issue).

  Since the method already filed by the applicant of the present application is a method applied to an internal combustion engine in which exhaust gas is not recirculated in the cylinder, it cannot be directly applied to the internal combustion engine with the EGR device of the present embodiment. However, since it will be helpful in understanding a method to be described later that can be applied to the present embodiment, a method already filed by the applicant will be described.

  The method already filed by the applicant of the present application is derived from the model shown in FIG. 2 (hereinafter referred to as a cylinder inflow gas amount model).

  The cylinder inflow gas amount model M20 shown in FIG. 2 includes a throttle model M21, an intake pipe model M22, and an intake valve model M23.

The throttle model M21, the opening degree (throttle opening) theta _t of the throttle valve 18 detected by a throttle opening sensor 43, the atmospheric pressure P _a, which is detected by the atmospheric pressure sensor 45, detected by the atmospheric temperature sensor 44 and the atmospheric temperature T _a, which is the pressure in the intake tube 13 calculated in the intake pipe model M22 (hereinafter, the intake pipe referred to as pressure) and P _m are inputted, in the model M21, from these input parameters , the flow rate of air passing through the throttle valve 18 per unit time (hereinafter referred to as the throttle valve passing air flow rate) m _t is calculated.

The intake pipe model M22, the throttle valve passage air flow rate m _t calculated at the throttle model M21 as described above, the flow rate of the gas flowing into the cylinder 5 to the per unit time calculated at the intake valve model M23 (hereinafter , referred to as cylinder inflow gas flow rate) and m _c, are input and the atmospheric temperature T _a, in the model M22, from these input parameters, the intake pipe pressure P _m, the temperature of the gas in the intake branch pipe 13 _Tm (hereinafter referred to as intake pipe temperature) is calculated.

The intake valve model M23, the intake pipe pressure P _m calculated at the intake pipe model M22 described above, the intake pipe temperature T _m, is input and the atmospheric temperature T _a, in the model M23, are those input The in-cylinder inflow gas flow rate _mc is calculated from the obtained parameters.

Then, in this method, as described later, the amount of gas flowing by utilizing the cylinder inflow gas flow rate m _c into the cylinder 5 (hereinafter, referred to as in-cylinder charged gas amount) M _c is calculated.

As can be seen from FIG. 2, in the in-cylinder inflow gas amount model M20, the parameter value calculated in each model is used as a parameter value input to another model. parameter value is input, the throttle opening theta _t, the atmospheric pressure P _a, and is only three parameters of the atmospheric temperature T _a. That is, according to this method, it can be said that the in-cylinder charged gas amount _Mc is calculated from the three parameters.

  Next, each of the models M21 to M23 will be described in detail.

In the throttle model M21, enter the atmospheric pressure P _a and the atmospheric temperature T _a and the intake pipe pressure P _m and the throttle opening theta _t in equation (1), by solving this equation, the throttle valve passage air flow rate m _t is calculated.

In the equation (1), μ _t is a flow coefficient in the throttle valve, is a function of the throttle opening θ _t , and is determined from the map shown in FIG. Also, A _t is the cross-sectional area of the opening of the throttle valve, a function of the throttle opening theta _t, determined from the map shown in FIG. It may be determined from a single map these flow coefficient mu _t and the opening cross-sectional area A _t summarizes mu _t · A _t as a function of the throttle opening theta _t. Further, R is a constant related to a gas constant is a value obtained by dividing the mass M _a of air per 1 mole of the so-called gas constant ^{R * (R = R * /} M a)
Further, Φ (P _m / P _a ) is a function having P _m / P _a as a variable, as shown in the following equation (2).

  In the equation (2), κ is a specific heat ratio, and is a constant value in this method.

Note that there is a relationship as shown in FIG. 5 between the function Φ (P _m / P _a ) and P _m / P _a . Therefore, instead of the equation (2), a map for calculating _a function Φ (P _m / P _a ) having P _m / P _a as a variable is stored in the ROM 34 in advance, and P _m / P _a and this map From the above, the value of the function Φ (P _m / P _a ) may be calculated.

Note that these equations (1) and (2), the air pressure of the throttle valve 18 upstream of the air and the atmospheric pressure P _a, the temperature of the throttle valve 18 upstream of the air and the atmospheric temperature T _a, passes through the throttle valve 18 6 is the intake pipe pressure P _m , and the throttle valve 18 is based on the model as shown in FIG. 6, the law of conservation of mass, energy between the air upstream of the throttle valve 18 and the air that has passed through the throttle valve 18. Using the conservation equation and the relational expression established in the momentum conservation law, the equation of state of gas, the specific expression of the specific heat ratio (κ = Cp / Cv), and the Meyer's relational expression (Cp = Cv + R ^* ) Derived. Here, Cp is a constant pressure specific heat, Cv is a quantitative specific heat, and R ^* is a so-called gas constant.

Next, the intake pipe model M22 will be described. In the intake pipe model M22, by inputs the throttle valve passage air flow rate m _t and the cylinder inflow gas flow rate m _c and the atmospheric temperature T _a to the following equation (3) and (4), solving these equations, the intake pipe The pressure P _m and the intake pipe temperature T _m are calculated.

  In equations (3) and (4), V is the intake pipe 15 from the throttle valve 18 to the intake valve 6, the surge tank 14, the intake branch pipe 13, and the intake port 7 (hereinafter collectively referred to as the intake pipe portion and Generally), and is a constant value.

  These equations (3) and (4) are related to the intake pipe portion, and the air flowing into the intake pipe portion and the air flowing out from the intake pipe portion and flowing into the cylinder are based on the model shown in FIG. It is derived from the relational expression that holds in the law of conservation of mass and the law of conservation of energy.

Specifically, if the total air amount in the intake pipe portion is M, the temporal change in the total air amount M is the flow rate of air flowing into the intake pipe portion (that is, the throttle valve passage air flow rate) m _t. and the flow rate of air flowing into the outflow to the cylinder from the intake pipe portion (i.e., the cylinder inflow gas flow rate) to be equal to the difference between m _c, the mass conservation law, the following equation (5) is satisfied.

Then, the equation (3) is derived from the equation (5) and the gas equation of state ( _Pm · V = M · R ^* · T _m ).

Further, the temporal variation amount of energy M · Cv · T _m of the air in the intake pipe portion, the energy of the air entering the outlet to the cylinder from the energy amount and the intake pipe of the air flowing into the intake pipe portion since equal to the difference between the amount and the temperature of the air flowing into the intake pipe portion and the atmospheric temperature T _a, the temperature of the air flowing into the cylinder flows out from the intake pipe portion and the intake pipe temperature T _m, the energy The following equation (6) is established on the conservation law.

  Then, the equation (4) is derived from the equation (6) and the above-described gas state equation.

Next, the intake valve model M23 will be described. In the intake valve model M23, the intake pipe pressure P _m and the intake pipe temperature T _m and the atmospheric temperature T _a is input to the following equation (7), by solving this equation, the cylinder inflow gas flow rate m _c is calculated The

  In Expression (7), a and b are values obtained using the engine speed NE as a variable. In the internal combustion engine, when at least one of the valve timing corresponding to the valve opening timing or the valve closing timing of the intake valve 6 and the operating angle corresponding to the valve opening time can be changed, the equation (7) ), A and b are values obtained by using the engine speed NE and the valve timing and / or the phase angle as variables.

  In other words, in Equation (7), a is a proportional coefficient, and b is a value representing the amount of gas remaining in the cylinder 5 when the exhaust valve 8 is closed.

In the formula (7), when the engine operating condition has changed, i.e., at the time of transient operation, since the intake pipe temperature T _m may change significantly, to compensate for changes in the intake pipe temperature T _m as a correction coefficient for, it is used T _a / T _m derived from theory and empirical rules.

Note that equation (7) relates to the intake valve 6, and the basis of the model as shown in FIG. 8, to be proportional to the cylinder inflow gas flow rate m _c to the intake pipe pressure P _m as described in detail below Deemed and derived from theory and empirical rules.

That is, the in-cylinder charged gas amount M _c is proportional to the pressure in the cylinder 5 at the time of closing of the valve closing finalized and the intake valve during 6 of the intake valve 6. Here, since the pressure in the cylinder 5 when the intake valve 6 is closed can be regarded as being equal to the pressure of the air upstream of the intake valve 6 (ie, the intake pipe pressure) P _m , the cylinder charge gas amount _Mc is the intake air. It can be approximated to be proportional to the tube pressure P _m .

On the other hand, the in-cylinder charged gas amount M _c is the air flow rate (cylinder inflow gas flow rate) m _c flowing into the cylinder 5 during the valve opening period of the intake valve 6 over the valve opening period of the intake valve 6. It is obtained by time integration. Namely, relationship that is between the in-cylinder charged gas amount M _c and the cylinder inflow gas flow rate m _c, the time integral value of the cylinder inflow gas flow rate m _c is cylinder charged gas amount M _c.

Thus, in-cylinder charged gas amount M _c is proportional to the intake pipe pressure P _m, and, between the in-cylinder charged gas amount M _c and the cylinder inflow gas flow rate m _c, the cylinder inflow gas flow rate m _c It can be regarded as the time integral value of from that there is a relationship that a cylinder charged gas amount M _c, also inflow gas flow rate m _c-cylinder proportional to the intake pipe pressure P _m.

Therefore, in this method, it is assumed that the in-cylinder inflow gas flow rate _mc is proportional to the intake pipe pressure P _m , and Equation (7) is derived from theory and empirical rules.

Note that the cylinder inflow gas flow rate m _c calculated by Equation (7) are the average value of the flow rate of air flowing out from the intake pipe portion per unit time, the cylinder inflow gas flow rate m _c, the internal combustion engine by applying a time divided by the number of cylinders of the time required for one cycle, the in-cylinder charged gas amount M _c in each cylinder 5 is calculated.

  Next, this will be described with reference to FIG. 9, taking an internal combustion engine having four cylinders as an example.

  In FIG. 9, the horizontal axis is the crank angle, and the vertical axis is the amount of air flowing from the intake pipe portion into the cylinder 5 per unit time. In the example shown in FIG. 9, the intake stroke is performed in the order of the first cylinder # 1, the third cylinder # 3, the fourth cylinder # 4, and the second cylinder # 2. When the intake stroke is performed in this way, the flow rate of air flowing into each cylinder 5 from the intake pipe portion changes as indicated by the broken line in FIG. 9, and as a result, the flow rate of air flowing out from the intake pipe portion is 9 will change as indicated by the solid line in FIG.

The average value of the flow rate of air flowing out from the intake pipe portion (solid line in FIG. 9) is the in-cylinder inflow gas flow rate _mc , and is indicated by a broken line in FIG. Therefore, the cylinder charge gas amount M _c in each cylinder 5 is equal to the cylinder inflow gas flow rate m _c (broken line in FIG. 9) and the time taken for one cycle of the internal combustion engine (in the example shown in FIG. 9, the crankshaft is Time taken to rotate 720 °) divided by the number of cylinders (four in the example shown in FIG. 9), that is, time taken for the crankshaft to rotate 180 ° in the example shown in FIG. It is calculated by applying. Then, the cylinder charged gas amount M _c in each cylinder 5 calculated by Thus, for example, will correspond to the hatched in FIG.

Next, a method of calculating the cylinder charged gas amount M _c when the cylinder inflow gas amount model M20 described above is mounted on an internal combustion engine.

The in-cylinder charged gas amount _Mc is obtained from the equations (1) to (4) and (7) of each model of the in-cylinder inflow gas amount model M20. These five equations are used when mounted on an internal combustion engine. , It is discretized so that it can be processed by the ECU 31. That is, assuming that the time is _t and the calculation interval (calculation period) is Δt, these five expressions are discretized into the following expressions (8) to (12).

According to the in-cylinder inflow gas amount model M20 discretized and mounted in the internal combustion engine, the throttle valve passing air flow rate m _t (t) at time t calculated in the throttle model M21 and the intake valve model M23 The in-cylinder inflow gas flow rate m _c (t) at time t and the intake pipe temperature T _m (t) calculated at time t are input to the equations (10) and (11) of the intake pipe model M22. By solving Expressions (10) and (11), the intake pipe pressure P _m (t + Δt) and the intake pipe temperature T _m (t + Δt) at time (t + Δt) are calculated.

Then, the intake pipe pressure P _m (t + Δt) calculated in the intake pipe model M22 and the throttle opening θ _t (t) at the time t are input to the equations (8) and (9) of the throttle model M21. Is calculated, the throttle valve passing air flow rate m _t (t + Δt) at time (t + Δt) is calculated.

Further, the intake pipe pressure P _m (t + Δt) and the intake pipe temperature T _m (t + Δt) calculated in the intake pipe model M22 are input to the expression (12) of the intake valve model M23, and the time ( A cylinder inflow gas flow rate m _c (t + Δt) at t + Δt) is calculated.

By repeating such calculations, the cylinder inflow gas flow rate m _c at any time is calculated. Then, the in-cylinder inflow gas flow rate m _c thus calculated is multiplied by the time obtained by dividing the time taken for one cycle by the number of cylinders as described above, thereby obtaining the in-cylinder charged gas amount M of each cylinder at an arbitrary time. _c is calculated.

Incidentally, when the internal combustion engine is started, i.e., at time t = 0, the intake pipe pressure P _m is equal to the atmospheric pressure _{P a (P m (0)} = P a), whereas, the intake pipe temperature T _m It is set equal to the atmospheric temperature T _a (T _m (0) = T _a ), and calculation in each model M21 to M23 is started.

Further, as the atmospheric pressure P _a and the atmospheric temperature T _a is used in the cylinder inflow gas amount model M20 as described above, to always be using an atmospheric pressure and the atmospheric temperature when the calculation of the model M20 is started, Atmospheric pressure P _a (t) and atmospheric temperature T _a (t) at time _t may be used.

  Next, a cylinder inflow gas amount model applicable to the internal combustion engine with the EGR device shown in FIG. 1 will be described.

When referred to the flow rate of the cylinder inflow fresh air flow rate m _c-air air in the gas flowing into the cylinder (fresh air), the cylinder inflow fresh air flow rate m _c-air is found from the following equation (13).

In Expression (13), m _c is a cylinder inflow gas flow rate that is a flow rate of gas flowing into the cylinder, and m _c-egr is a cylinder inflow EGR that is a flow rate of exhaust gas in the gas flowing into the cylinder. The gas flow rate.

Here, when the engine is in steady operation (for example, when the throttle opening, EGR opening, engine speed, etc. are maintained substantially constant), the flow rate of exhaust gas passing through the EGR control valve 22 (hereinafter referred to as EGR). _{M egr (} referred to as control valve passage gas flow rate) and in-cylinder inflow EGR gas flow rate m _c-egr are equal ( _mc-egr = m _egr ).

  For this reason, Formula (13) can be rewritten to following Formula (14).

Therefore, Motomare and the cylinder inflow gas flow rate m _c and the EGR control valve passage gas flow rate m _egr, Motomari is the cylinder inflow fresh air flow rate m _c-air, therefore, the cylinder 5 at the time of closing of the intake valves 6 An in-cylinder charged fresh air amount _Mc-air, which is the amount of air (fresh air) filled, is obtained.

Therefore, in the first embodiment, as the cylinder inflow gas amount model for calculating the cylinder inflow gas flow rate m _c, using a model M10 shown in FIG. 10.

  The cylinder inflow gas amount model M10 shown in FIG. 10 includes a throttle model M11, an intake pipe model M12, an intake valve model M13, and an EGR control valve model M15.

The throttle model M11 is the same model as the throttle model M21 shown in FIG. 2, and the intake valve model M13 is also the same model as the intake valve model M23 shown in FIG. to describe in, in the throttle model M11, and inputs the throttle opening theta _t and the atmospheric pressure P _a and the atmospheric temperature T _a and the intake pipe pressure P _m in the formula (1), by solving this equation, it passes the throttle valve calculated air flow rate m _t is the intake valve model M13, inputs the atmospheric temperature T _a and the intake pipe pressure P _m and the intake pipe temperature T _m in the formula (7), by solving this equation, the cylinder inflow A gas flow rate _mc is calculated.

The intake pipe model M12, the throttle valve passage air flow rate m _t calculated at the throttle model M11, the cylinder inflow gas flow rate m _c calculated at the intake valve model M13, the atmospheric temperature T _a, EGR control valve model A flow rate of exhaust gas passing through the EGR control valve 22 per unit time calculated in M15 (hereinafter referred to as an EGR control valve passage gas flow rate) m _egr is input, and in the model M12, from these input parameters The intake pipe pressure P _m and the intake pipe temperature T _m are calculated.

Further, the EGR control valve model M15 is an EGR opening degree theta _e and the atmospheric pressure P _a and the atmospheric temperature T _a and the intake pipe pressure P _m calculated at the intake pipe model M12 and the exhaust temperature T _e is input, the In the model M15, the EGR control valve passage gas flow rate m _egr is calculated from these input parameters.

As can be seen from FIG. 10, in the in-cylinder inflow gas amount model M10, the parameter value calculated in each model is used as a parameter value input to another model. parameter value is input, the throttle opening theta _t, EGR opening degree theta _e, the atmospheric pressure P _a, and is only four parameters of the atmospheric temperature T _a. That is, according to this embodiment, it can be said that the cylinder charge gas amount _Mc is calculated from these four parameters.

  Next, the intake pipe model M12 and the EGR control valve model M15 will be described in detail.

In the intake pipe model M12, the throttle valve passage air flow rate m _t and the cylinder inflow gas flow rate m _c and the atmospheric temperature T _a and the EGR control valve passage gas flow rate m _egr by the following equation and an exhaust temperature T _e (15) and (16 ) And solving these equations, the intake pipe pressure P _m and the intake pipe temperature T _m are calculated.

  Also in the equations (15) and (16), V is the volume of the intake pipe portion from the throttle valve 18 to the intake valve 6 and is usually a constant value.

  As described above with respect to the intake pipe model M22, these equations (15) and (16) relate to the intake pipe portion and flow into the intake pipe portion based on the model shown in FIG. It is derived from a relational expression that is established based on the law of conservation of mass and the law of conservation of energy among the air that flows, the exhaust gas that flows into the intake pipe portion, and the gas that flows out of the intake pipe portion and flows into the cylinder.

Further, the EGR control valve model M15, by inputting the EGR opening theta _e and the intake pipe pressure P _m and the exhaust pressure P _e and the exhaust temperature T _e in the following equation (17), the EGR control valve passage gas flow rate m _egr is calculated.

In Expression (17), μ _e is a flow coefficient in the EGR control valve 22 and is a function of the EGR opening degree θ _e . A _e is the opening cross-sectional area of the EGR control valve 22 and is a function of the EGR opening degree θ _e . Further, R is a constant related to a gas constant is a value obtained by dividing the mass M _e of the exhaust gas per 1 mol of a so-called gas constant ^{_{R * (R e = R *}} / M e).

_Pe is the pressure of the exhaust gas upstream of the EGR control valve 22, and _Te is the temperature of the exhaust gas upstream of the EGR control valve 22.

Further, Φ (P _m / P _e ) is a function having P _m / P _e as a variable, as shown in the following equation (18).

This equation (18) is obtained by replacing the variable P _m / P _a in the equation (2) with the variable P _m / P _e . Therefore, κ is a specific heat ratio, and is a constant value in this embodiment.

Also, there is a relationship as shown in FIG. 5 between the function Φ (P _m / P _e ) and P _m / P _e . Therefore, instead of the equation (18), a map for calculating the function Φ (P _m / P _e ) with P _m / P _e as a variable is stored in the ROM 34 in advance, and P _m / P _e and this map From the above, the value of the function Φ (P _m / P _e ) may be calculated.

Note that these equations (17) and (18), the pressure of the EGR control valve 22 upstream of the exhaust gas and the exhaust pressure P _e, the temperature of the EGR control valve 22 upstream of the exhaust gas and the exhaust temperature T _e, the EGR control valve as an intake pipe pressure P _m and the pressure of the exhaust gas passing through the 22 relates EGR control valve, in the basic model shown in FIG. 12, has passed through the EGR control valve 22 upstream of the exhaust gas and the EGR control valve 22 It is derived by using the relational expression that is established with respect to the exhaust gas in terms of the law of conservation of mass, the law of conservation of energy, and the law of conservation of momentum, as well as the equation of state of gas, the definition of specific heat ratio, and the Meyer's relational expression. .

  That is, the basic idea for deriving these equations (17) and (18) is the same as the idea for deriving equations (1) and (2) for calculating the throttle valve passage air flow rate. It is.

Thus, according to this embodiment, by the high pressure P _e and the exhaust temperature T _e and the intake pipe pressure P _m and the EGR opening theta _e entered in the formula and numerical calculation, EGR control valve passage Since the gas flow rate is calculated and therefore the number of maps used at this time is reduced, the map creation effort can be greatly reduced.

Next, a method for calculating the in _- cylinder charged fresh air amount _Mc-air when the in-cylinder inflow gas amount model M10 is mounted on the internal combustion engine will be described.

The in-cylinder charged fresh air amount _Mc-air is obtained from the equations (1), (2), (7), and (15) to (18) of each model of the in-cylinder inflow gas amount model M10. Is mounted so that it can be processed by the ECU 31 when mounted on an internal combustion engine. That is, when time is t and calculation interval (calculation cycle) is Δt, equations (1), (2), and (7) are discretized into equations (8), (9), and (12), respectively. .

  Moreover, Formula (15)-(18) is discretized by following Formula (19)-(22), respectively.

Further, equations for calculating the exhaust pressure _Pe and the exhaust temperature _Te are also discretized as the following equations (23) and (24). Details regarding these equations will be described later.

According to the cylinder inflow gas amount model M10 discretized and mounted in the internal combustion engine in this way, the throttle valve passing air flow rate m _t (t) at the time t calculated in the throttle model M11 and the EGR control valve model EGR control valve passage gas amount m _egr (t) at time t calculated in M15, in-cylinder inflow gas flow rate m _c (t) at time t calculated in intake valve model M13, and exhaust temperature T at time _{t e} (t) is input to the equations (19) and (20) of the intake pipe model M12, and by solving these expressions, the intake pipe pressure P _m (t + Δt) and the intake pipe temperature T _m (at time (t + Δt)) t + Δt) is calculated.

Then, the intake pipe pressure P _m (t + Δt) at the time (t + Δt) calculated in the intake pipe model M12 and the throttle opening θ _t (t + Δt) at the same time (t + Δt) are expressed by the equation (8) of the throttle model M11 and By inputting into (9) and solving these equations, the throttle valve passing air flow rate m _t (t + Δt) at time (t + Δt) is calculated.

Further, the intake pipe pressure P _m (t + Δt) and the intake pipe temperature T _m (t + Δt) at time (t + Δt) calculated in the intake pipe model M12 are input to the expression (12) of the intake valve model M13, and this expression is solved. Thus, the in-cylinder inflow gas flow rate m _c (t + Δt) at the time (t + Δt) is calculated.

The cylinder inflow gas flow rate m _c (t + Δt) at time (t + Δt) calculated in the intake pipe model M12 and the EGR control valve passage gas flow rate m at time t calculated in the EGR control valve model M15 in the previous routine. _By entering _egr (t) into equation (14) and solving this equation, the cylinder inflow fresh air flow rate _mc-air (t + Δt) at time (t + Δt) is calculated.

The intake pipe pressure P _m (t + Δt) at the time (t + Δt) calculated in the intake pipe model M12, the exhaust pressure P _e (t + Δt) at the same time (t + Δt), and the EGR opening θ at the same time (t + Δt) are calculated. _e (t + Δt) is input to equations (21) and (22) of the EGR control valve model M15, and by solving these equations, the EGR control valve passage gas flow rate m _egr (t + Δt) at time (t + Δt) is calculated. The

Further, the engine load factor KL (t + Δt) at time (t + Δt) and the engine speed NE at the same time (t + Δt) are input to equation (23), and the exhaust pressure at time (t + Δt) is solved by solving this equation. P _e (t + Δt) is calculated.

Further, the EGR control valve passage gas flow rate m _egr (t) at time t and the engine speed NE (t + Δt) at time (t + Δt) are input to equation (24), and by solving this equation, time (t + Δt) exhaust temperature T _e (t + Δt) is calculated in).

The thus calculated m _c (t + Δt), m _egr (t + Δt), and m _t (t + Δt) are input again to the equations (19) and (20) of the intake pipe model M12.

By repeating such calculation, the in-cylinder fresh air flow rate _mc-air at the time is arbitrarily calculated. Then, the in-cylinder inflow fresh air flow rate _mc-air calculated in this way is multiplied by the time obtained by dividing the time taken for one cycle by the number of cylinders as described above, so that the in-cylinder filling of each cylinder at an arbitrary time is performed. A fresh air amount Mc _-air is calculated.

Incidentally, when the internal combustion engine is started, i.e., at time t = 0, the intake pipe pressure P _m is equal to the atmospheric pressure _{P a (P m (0)} = P a), the intake pipe temperature T _m and exhaust temperature T _e is equal to the atmospheric temperature _{T a (T m (0)} = T a, T e (0) = T a), is calculated for each model M11~M13 is started.

Further, as the atmospheric pressure P _a and the atmospheric temperature T _a is used in the cylinder inflow gas amount model M10 as described above, to always be using an atmospheric pressure and the atmospheric temperature when the calculation of the model M10 is started, Atmospheric pressure P _a (t) and atmospheric temperature T _a (t) at time t may be used.

Further, in the present embodiment, the exhaust pressure P _e, the exhaust temperature T _e, the intake pipe pressure P _m, and, from equation (17) utilizing the four parameters of the EGR opening degree theta _e, EGR control valve passage gas flow rate m _egr is calculated, but these four parameters are used, but the EGR control valve passage gas flow rate m _egr may be obtained from an equation different from equation (17), or these four parameters are used as arguments. The EGR control valve passage gas flow rate m _egr may be obtained from the map.

  Further, according to this embodiment, since the EGR control valve passage gas flow rate is calculated with high accuracy, the in-cylinder charged fresh air amount calculated based on the EGR control valve passage gas flow rate is also calculated with high accuracy. The fuel injection amount for making the air-fuel ratio the target air-fuel ratio is calculated with high accuracy.

  Furthermore, the EGR opening degree may be feedback-controlled using the EGR control valve passage gas flow rate calculated according to the present embodiment so that the EGR control valve passage gas amount becomes the target value.

  Next, a method for deriving Equation (23) for calculating the exhaust pressure will be described.

  The engine load factor (%) is the ratio of the amount of gas (g) actually charged in the cylinder to the amount of gas (g) that occupies the maximum volume of the cylinder in the standard state. Obtained from

In Expression (25), _Mc-air is the in _- cylinder charged fresh air amount (g) that is the amount of fresh air that is filled in each cylinder when the intake valve is closed, and DSP is the amount of fresh air in the internal combustion engine. It is the displacement (liter), NCYL is the number of cylinders, and ρ _astd is the air density (about 1.2 g / liter) in the standard state (1 atm, 25 ° C.).

Referring to Figure 13 showing the relationship between the engine load factor KL and the engine speed NE and the exhaust pressure P _e, the higher the the exhaust pressure P _e engine load factor KL increases, the exhaust pressure engine speed NE increases P _e is made it can be seen that high. Then, the exhaust pressure P _e is primarily because highly dependent on the these engine load factor KL and the engine speed NE, the accuracy of the exhaust pressure obtained from these engine load factor KL and the engine speed NE is relatively high.

Therefore, the exhaust pressure P _e, as in equation (23) can be expressed by a function f ₁ that the engine load factor KL and the engine speed NE as variables (KL, NE). In the first embodiment, this function f ₁ (KL, NE) is stored in advance in the ROM 34 in the form of a map with the engine load factor KL and the engine speed NE as variables, and the engine load factor KL and the engine speed NE are stored. And the exhaust pressure _Pe is calculated from this map.

According to this, since it is not necessary to provide a sensor for detecting the exhaust pressure in the internal combustion engine, it is possible to accurately detect the exhaust pressure while suppressing the cost of the internal combustion engine, and in turn, it passes through the EGR control valve. The gas flow rate m _egr can be calculated with high accuracy.

  Also, in an internal combustion engine with an EGR device, when detecting exhaust pressure using a map without using a sensor, the engine speed, fuel injection amount, fuel ignition timing, EGR rate (filled in the cylinder) A map having as arguments a number of parameters such as the ratio of the amount of exhaust gas (g) to the amount of gas (g)) must be used. However, since the engine load factor and the engine speed can be used as parameters by using the engine load factor as a parameter as in this embodiment, the map creation effort can be reduced according to this embodiment. Can be reduced.

  Of course, if there is no problem with increasing the cost of the internal combustion engine, the exhaust pressure may be detected using a sensor.

Further, in the case of calculating the exhaust pressure from the function f ₁ described above, the cylinder charge fresh air flow rate _mc-air may be used as a variable instead of the engine speed NE.

Next, a method of deriving the equation for calculating the exhaust temperature T _e (24).

When the amount of heat generated by the exhaust gas newly flowing into the EGR passage 21 is referred to as input heat amount, and the amount of heat released from the EGR passage 21 to the atmosphere is referred to as heat release amount, the EGR control valve passage gas flow rate m _egr increases. When the amount of exhaust gas flowing into the EGR passage 21 increases, the amount of input heat becomes larger than the amount of heat released. Therefore, the exhaust gas temperature increases as the EGR control valve passage gas flow rate m _egr increases.

  Further, when the engine speed increases, the temperature of the exhaust gas discharged from the cylinder itself increases.

In other words, the relationship between the EGR control valve passage gas flow rate m _egr and the engine speed NE and the exhaust temperature T _e is as shown in FIG. 14.

Then, the exhaust temperature T _e is mainly because it largely depends on the these EGR control valve passage gas flow rate m _egr and the engine speed NE, these EGR control valve passage gas flow rate m _egr and the engine speed NE The accuracy of the exhaust temperature required is relatively high.

Therefore, the exhaust temperature T _e, as in equation (24), expressed by the EGR control valve passage gas flow rate m _egr and the engine speed function has a NE as a variable f ₂ (m _egr, NE). In the first embodiment, this function f ₂ (m _egr , NE) is stored in advance in the ROM 34 in the form of a map using the EGR control valve passage gas flow rate m _egr and the engine speed NE as variables, and passed through the EGR control valve. The exhaust gas temperature _Te is calculated from the gas flow rate m _egr , the engine speed NE, and this map.

According to this, since it is not necessary to provide a sensor for detecting the exhaust gas temperature in the internal combustion engine, it is possible to detect the exhaust gas temperature accurately while suppressing the cost of the internal combustion engine. The gas flow rate m _egr can be calculated with high accuracy.

  Of course, if the increase in the cost of the internal combustion engine is not a problem, the exhaust temperature may be detected using a sensor.

  Further, as described above, since the temperature of the exhaust gas discharged from the cylinder changes according to the engine speed NE, the temperature of the exhaust gas can be obtained from a one-dimensional map using the engine speed NE as an argument. .

However, when calculating the EGR control valve passage gas flow rate m _egr at the EGR control valve model M15, the exhaust temperature T _e, rather than using the temperature of the exhaust gas discharged from the cylinders, the EGR control valve 22 upstream of the exhaust gas The EGR control valve passage gas flow rate m _egr can be calculated more accurately by using the temperature.

According to this embodiment, as the exhaust temperature T _e used to calculate the EGR control valve passage gas flow rate m _egr, it means using a temperature of the EGR control valve 22 upstream of the exhaust gas, the EGR control valve passage gas The flow rate m _egr can be calculated with high accuracy.

In addition, when calculating the exhaust gas temperature from the function f ₂ described above, the in-cylinder charged fresh air flow rate _mc-air may be used as a variable instead of the engine speed NE.

  By the way, in the above-described embodiment, it is assumed that the operation is steady, and therefore, the flow rate of exhaust gas flowing into the intake pipe portion (hereinafter referred to as inflowing exhaust gas partial flow rate) is equal to the EGR control valve passage gas flow rate. The flow rate of exhaust gas flowing into the cylinder (hereinafter referred to as in-cylinder inflow exhaust gas flow rate) is also assumed to be equal to the EGR control valve passage gas flow rate.

  However, during transient operation, for example, when the EGR opening changes and the EGR control valve passage gas flow rate changes, the inflow exhaust gas partial flow rate and the cylinder inflow exhaust gas flow rate are temporarily set to the EGR control valve passage gas flow rate. Are not equal. That is, these inflow exhaust gas partial flow rate and in-cylinder inflow exhaust gas flow rate follow the change in the EGR control valve passage gas flow rate but change with a delay.

  Therefore, in order to maintain high accuracy of the parameter values calculated in the above-described embodiment during the transient operation, the inflow exhaust gas partial flow rate and the inflow exhaust gas flow rate in the cylinder change in the flow rate of the EGR control valve passage gas. It is necessary to consider that it changes late.

Considering the inflow exhaust gas partial flow rate when the EGR control valve passage gas flow rate changes, as shown in FIG. 15, the inflow exhaust gas partial flow rate m _egr-egr is _equal to the EGR control valve passage gas flow rate m _egr. It does not change until a certain time Td elapses after the change, and changes toward the EGR control valve passage gas flow rate m _egr with a primary delay after the certain time Td elapses.

  Here, the fixed time Td is a time required for the exhaust gas that has passed through the EGR control valve 22 to reach the intake pipe portion, and the higher the flow rate of the exhaust gas that has passed through the EGR control valve 22, the more the intake air accordingly. The lower the tube pressure, and hence the shorter the engine speed, the shorter. That is, the fixed time (hereinafter referred to as dead time) Td is a function of the engine speed as shown in FIG.

  On the other hand, the time constant τ of the first-order lag represents the ease of diffusion of the exhaust gas that has passed through the EGR control valve 22, and the higher the flow rate of the exhaust gas that has passed through the EGR control valve 22, the higher the intake pipe. The lower the pressure, the smaller the engine speed. That is, the time constant τ is also a function of the engine speed as shown in FIG.

  Therefore, in the present embodiment, the dead time Td1 is stored in advance in the ROM 34 in the form of a map as shown in FIG. 16A, and when the EGR control valve passage gas flow rate changes, it is based on the engine speed NE. The dead time Td1 is calculated, and the inflow exhaust gas partial flow rate is calculated on the assumption that the inflow exhaust gas partial flow rate starts to change toward the EGR control valve passage gas flow rate after the dead time Td1 has elapsed.

  In this embodiment, the time constant τ1 is stored in advance in the ROM 34 in the form of a map as shown in FIG. 16B, and when the EGR control valve passage gas flow rate changes, it is based on the engine speed NE. The time constant τ1 is calculated, and the inflowing exhaust gas partial flow rate is calculated on the assumption that the inflowing exhaust gas partial flow rate changes toward the EGR control valve passage gas flow rate with the time constant τ1 after the dead time Td1 has elapsed. .

  The concept of dead time and first-order lag concerning the inflowing exhaust gas partial flow rate is similarly applied to the inflowing exhaust gas flow rate in the cylinder.

  That is, the dead time Td2 relating to the in-cylinder inflow exhaust gas flow rate is previously stored in the ROM 34 in the form of a map as shown in FIG. 17A, and when the EGR control valve passage gas flow rate changes, the engine speed NE The in-cylinder inflow exhaust gas flow rate is calculated on the assumption that the in-cylinder inflow exhaust gas flow rate starts to change toward the EGR control valve passage gas flow rate after the dead time Td2 has elapsed. To do.

  Then, the time constant τ2 regarding the in-cylinder inflow exhaust gas flow rate is stored in advance in the ROM 34 in the form of a map as shown in FIG. 17B, and when the EGR control valve passage gas flow rate changes, the engine speed NE The time constant τ2 is calculated based on the above, and the cylinder inflow exhaust gas flow rate is assumed to change toward the EGR control valve passing gas flow rate with the time constant τ2 after the dead time Td2 has elapsed. Calculate the flow rate.

  Since the distance from the EGR control valve 22 to the cylinder is longer than the distance from the EGR control valve 22 to the intake pipe portion, the dead time Td2 related to the in-cylinder inflow exhaust gas flow rate is greater than the dead time Td1 related to the inflow exhaust gas partial flow rate. The time constant τ2 related to the in-cylinder inflow exhaust gas flow rate tends to be larger than the time constant τ1 related to the inflow exhaust gas partial flow rate.

By the way, in the first embodiment, the EGR control valve passage gas flow rate m _egr is calculated by using a calculation formula, that is, the formulas (17) and (18). Even if not, a method of calculating the EGR control valve passage gas flow rate m _egr relatively easily was devised. Therefore, instead of using the equations (17) and (18), the EGR control valve passage gas flow rate may be calculated using the method described below.

Cylinder charged gas amount M _c is known to be expressed by a linear function of the intake pipe pressure P _m of when the intake valve 6 is closed. That is, according to theory and empirical rule, the in-cylinder charged gas amount _Mc is proportional to the in-cylinder pressure when the intake valve 6 is closed, and this in-cylinder pressure is the mixed gas pressure upstream of the intake valve 6, that is, It almost coincides with the intake pipe pressure P _m .

When EGR gas is not supplied, only fresh air is filled into the cylinder, and therefore, the cylinder fresh charge amount M _{c-air at} this time, and therefore the engine load factor KL, is expressed as a linear function equation of the intake pipe pressure P _m. Can be expressed as That is, the engine load factor KL can be obtained easily and accurately.

However, the situation is completely different when EGR gas is supplied, and not only fresh air but also EGR gas is filled in the cylinder. Therefore, in the conventional in-cylinder charging fresh air amount M _c-air or, is the fact was thought absolutely impossible to represent the engine load factor KL by a linear function expression of the intake pipe pressure P _m.

If it is possible to represent the in-cylinder charged EGR gas amount M _c-egr by a linear function of the intake pipe pressure P _m, it can represent a cylinder charged gas amount M _c by a linear function of the intake pipe pressure P _m , considering that cylinder charged gas amount M _c is the sum of the cylinder charging fresh air amount M _c-air and cylinder charged EGR gas amount M _c-egr, cylinder when the EGR gas is being supplied The internal charge fresh air amount M _c-air or the engine load factor KL can be expressed by a linear function expression of the intake pipe pressure P _m .

However, in the conventional, cylinder charged EGR gas amount M _c-egr is also had been considered can not be represented by a linear function expression of the intake pipe pressure P _m. This will be described with reference to FIGS.

First, as described above, the EGR control valve passage gas flow rate m _egr (g / sec) is expressed by Expression (17), and the function Φ (P _m / P _e ) is expressed by Expression (18).

Here, the exhaust pressure P _e in order to simplify the calculation when the atmospheric pressure P _a, EGR control valve passage gas flow rate m _egr represented by the formula (17) is as shown in FIG 18. That is, the EGR control valve passage gas flow rate m _egr is maintained substantially constant when the intake pipe pressure P _m is small, and when the intake pipe pressure P _m becomes high, it is nonlinear with respect to the intake pipe pressure P _m as indicated by NR in FIG. It decreases toward the atmospheric pressure P _a while showing sex. Note that the nonlinear portion NR is due to P _e / √T _e parts and functions Φ (P _m / P _e) of the formula (17).

Therefore, it has been considered that the EGR control valve passage gas flow rate m _egr , especially the nonlinear portion NR cannot be expressed by a linear function expression of the intake pipe pressure P _m . However, if a considerably large number of linear function equations are used, it is considered that the EGR control valve passage gas flow rate m _egr can be expressed by a linear function equation of the intake pipe pressure P _m . However, in this case, the engine load factor KL can no longer be simply obtained.

However, according to the inventors of the present application, the EGR control valve passage gas flow rate m _egr can be expressed by two linear function expressions of the intake pipe pressure P _m , and therefore, the in-cylinder charged fresh air amount M _c-air , or is the it has been found that it is possible to represent the engine load factor KL by two first-order function equation of the intake pipe pressure P _m.

That is, first, as shown in FIG. 19, the exhaust temperature T _e is to increase the intake pipe pressure P _m, the exhaust pressure P _e increases significantly than increases, as a result, P _e / √T _e the it can be expressed by a linear function expression of the intake pipe pressure P _m.

The function Φ (P _m / P _e ) can also be expressed by a linear function expression of the intake pipe pressure P _m . This will be described with reference to FIG. Considering that the exhaust pressure Pe is not maintained at _a constant atmospheric pressure _Pa but varies according to the intake pipe pressure P _m , as shown in FIG. 20A, the intake pipe pressure P _m The function Φ (P _m / P _e ) when is P _m 1 is not on the curve C _a that converges to the atmospheric pressure P _a , but is on the curve C 1 that converges to the exhaust pressure P _e 1, which is plotted (○) Similarly, Φ (P _m / P _e ) when P _m = P _m 2 (> P _m 1) is on the curve C2 that converges to the exhaust pressure P _e 2 (> P _e 1), and P _m = Φ (P _m / P _e ) when P _m 3 (> P _m 2) is on a curve C3 that converges to the exhaust pressure P _e 3 (> P _e 2).

The plot obtained in this way can be connected by a straight line L2, as shown in FIG. Therefore, the function Φ (P _m / P _e) is a linear function of the intake pipe pressure P _m corresponding to the straight line L1 when the intake pipe pressure P _m is small, corresponds to the straight line L2 when the intake pipe pressure P _m is greater intake can be represented by a linear function of the pipe pressure P _m, it will be able to thus to represent in two linear functions of the intake pipe pressure P _m. That is, the EGR control valve passage gas flow rate m _egr can be expressed by two linear function expressions of the intake pipe pressure P _m .

Here, during steady operation, the EGR control valve passage gas flow rate m _egr (g / sec) and the in-cylinder inflow EGR gas flow rate m _c-egr (g / sec) are equal to each other. The cylinder filling EGR gas amount M _c-egr is _{calculated by adding the} time required for one intake stroke of each cylinder to the cylinder inflow EGR gas flow rate m _c-egr , that is, the time required for one cycle of the internal combustion engine. It is obtained by multiplying the time ΔT (sec) divided by (M _c-egr = m _c-egr · ΔT).

Then, the cylinder filling EGR gas amount M _c-egr during steady operation can be expressed by a linear function expression of the intake pipe pressure P _m .

Accordingly, the in-cylinder charged fresh air amount M _c-air or the engine load factor KL during steady operation and when EGR gas is supplied can be expressed by two linear function expressions of the intake pipe pressure P _m. It turns out that.

Assuming that the engine load factor KL when the EGR gas is supplied is KLon, FIG. 21 shows that during steady operation (for example, when the engine speed NE and the EGR opening degree θ _e are kept constant). An example of two linear function expressions of the intake pipe pressure P _m representing the engine load factor KLon is shown. As shown in FIG. 21, the engine load factor KLon, the slope is continuous in different and the connection point CP with each other is represented by two first-order function equation of the intake pipe pressure P _m. That is, when the intake pipe pressure P _m is small, the engine load factor KLon is expressed by a linear function equation of the gradient e1, and when the intake pipe pressure P _m is high, the engine load factor KLon is expressed by a linear function equation of the gradient e2. .

  Here, assuming that the gradients of the two linear function equations are e1 and e2, respectively, and the intake pipe pressure and the engine load factor at the connection point CP are d and r, respectively, these two linear function equations can be expressed by the following equations. .

KLon = e1 · (P _m −d) + r... P _m ≦ d
KLon = e 2 · (P _m −d) + r... P _m > d
These are collectively expressed as the following equation (26).

KLon = e · (P _m −d) + r (26)
e = e1 ... P _m ≦ d
e = e2 ... P _m> d

In this embodiment, two linear functions of the intake pipe pressure P _m representing the engine load factor KLon at the time of steady state operation is stored in advance in the ROM34 in the form shown in equation (26). In this way, two linear function expressions can be represented by three parameters e, d, and r. That is, it is possible to reduce the number of parameters necessary to express two linear function expressions.

  Each parameter e, d, r of this formula (26) is calculated based on the following formula.

e1 = e1 ^* · ktha
e2 = e2 ^* · ktha
d = d ^* · ktha · kpa
r = r ^* · ktha · kpa
Here, e1 ^* , e2 ^* , d ^* , and r ^* are the gradient, the intake pipe pressure and the engine load factor at the connection point, respectively, when the engine ambient environment is a predetermined reference environment. Any state may be used as the reference environment state, but in this embodiment, the standard state (1 atm, 25 ° C.) is used as the reference environment state.

On the other hand, ktha represents an atmospheric temperature correction coefficient, and kpa represents an atmospheric pressure correction coefficient. The atmospheric temperature correction coefficient ktha is for correcting each parameter e1 ^* , e2 ^* , d ^* , r ^* in the reference environment state according to the actual atmospheric temperature detected by the atmospheric temperature sensor 44, and is corrected. When there is no need, 1.0 is set. The atmospheric pressure correction coefficient kpa is used to correct the parameters d ^* and r ^* in the reference environment state according to the actual atmospheric pressure detected by the atmospheric pressure sensor 45, and when correction is not necessary. 1.0.

Therefore, considering that the atmospheric temperature correction coefficient ktha or the atmospheric pressure correction coefficient kpa is a representative value representative of the actual engine ambient environment condition, the reference environment condition is based on the representative value representative of the actual engine environment environment condition. This means that the parameters e1 ^* , e2 ^* , d ^* , and r ^* are corrected. Alternatively, it can be considered that the engine load factor KLon in the reference environmental state is corrected based on the representative value representative of the actual environmental environment surrounding the engine.

On the other hand, in consideration of the fact that the opening cross-sectional area A _e of the EGR control valve 22 in equation (17) depends on the EGR opening θ _e and the engine charging efficiency depends on the engine speed NE, ^{e * (e1 *, e2 *} ), d *, and the r ^* is set according to the EGR opening theta _e or the engine speed NE.

Specifically, as shown in FIG. 22 (A), the gradient e1 ^* increases as the engine speed NE increases when the engine speed NE is low, and the engine speed NE when the engine speed NE is high. It becomes smaller as NE becomes higher, and further becomes larger as EGR opening degree θ _e becomes larger. Further, as shown in FIG. 22B, the gradient e2 ^* increases as the engine speed NE increases when the engine speed NE is low, and the engine speed NE increases when the engine speed NE is high. It becomes smaller as the EGR opening degree θ _e becomes larger. These gradients e1 ^*, e2 ^* are obtained by experiments in advance, advance in the ROM34 in the form of a map shown in FIG. 22 (C) and (D) as a function of the respective engine speed NE and EGR opening theta _e It is remembered.

On the other hand, the intake pipe pressure d ^{* at} the connection point CP decreases as the engine speed NE increases, as shown in FIG. The intake pipe pressure d ^{* at the} connection point CP is also obtained in advance by experiments, and is stored in advance in the ROM 34 in the form of a map shown in FIG. 23 as a function of the engine speed NE.

Further, as shown in FIG. 24A, the engine load factor r ^{* at} the connection point CP increases as the engine speed NE increases when the engine speed NE is low, and the engine load factor r ^* increases when the engine speed NE is high. It decreases as the rotational speed NE increases, and further decreases as the EGR opening degree θ _e increases. Has been determined by the engine load factor r ^* also advance by experiments at the connection point CP, is stored in advance in the ROM34 in the form of a map shown in FIG. 24 (B) as a function of the engine speed NE and the EGR opening theta _e Yes.

Therefore, generally speaking, with respect to a plurality of different EGR openings θ _e , the in-cylinder charged fresh air amount M _c-air during steady operation or the intake pipe pressure P _m representing the engine load factor KLon is 2 One linear function expression is obtained and stored in advance. Further, two linear function expressions of the in-cylinder charged fresh air amount M _c-air during steady operation or the intake pipe pressure P _m representing the engine load factor KLon are obtained in advance for a plurality of different engine speeds NE. It also means that it is remembered.

Figure 25 is in and various EGR opening theta _e at a constant engine speed NE, it shows an example of two linear functions of the intake pipe pressure P _m representing the engine load factor KLon at the time of steady operation. The broken line in FIG. 25 indicates the engine load factor KLoff when EGR gas is not supplied, that is, when the EGR opening degree θ _e is zero.

On the other hand, as described above, can represent the engine load factor KLoff when the EGR gas is not being supplied by a linear function expression of the intake pipe pressure P _m. Figure 26 is a steady operation timing (e.g., function speed NE is when it is held constant) are shown an example of two linear functions of the intake pipe pressure P _m representing the engine load factor KLoff of. In the present embodiment, as shown in FIG. 26, the engine load factor KLoff is the slope is continuous in different and the connection point CP with each other is represented by two first-order function equation of the intake pipe pressure P _m. That is, when the intake pipe pressure P _m is small, the engine load factor KLoff is expressed by a linear function equation of the gradient _a 1, and when the intake pipe pressure P _m is high, the engine load factor KLoff is expressed by a linear function equation of the gradient _a 2. .

  Here, assuming that the gradients of the two linear function equations are a1 and a2, respectively, and the intake pipe pressure and the engine load factor at the connection point CP are d and c, respectively, these two linear function equations can be expressed by the following equations. .

_{KLoff = a1 · (P m -d} ) + c ... P m ≦ d
KLoff = a 2 · (P _m −d) + c... P _m > d
These are collectively expressed as the following equation (27).

KLoff = a · (P _m −d) + c (27)
a = a1 ... P _m ≦ d
a = a2... P _m > d

In this embodiment, two linear functions of the intake pipe pressure P _m representing the engine load factor KLoff at the time of steady state operation is stored in advance in the ROM34 in the form shown in equation (27). In this case, the intake pipe pressure d at the connection point CP is the same as the intake pipe pressure d at the connection point CP for the engine load factor KLon described above. Therefore, the number of parameters can be further reduced. Of course, the intake pipe pressures at these connection points CP may be different from each other.

  The parameters a and r in the equation (27) are calculated based on the following equations.

a1 = a1 ^* · ktha
a2 = a2 ^* · ktha
c = c ^* · ktha · kpa
Here, a1 ^* , a2 ^* , and r ^* are the engine load factor at the gradient and the connection point, respectively, when the engine ambient environment state is the above-described reference environment state, that is, the standard state.

Therefore, considering that the atmospheric temperature correction coefficient ktha or the atmospheric pressure correction coefficient kpa is a representative value representative of the actual engine ambient environment condition, the reference environment condition is based on the representative value representative of the actual engine environment environment condition. In other words, the parameters a1 ^* , a2 ^* , and c ^* are corrected. Alternatively, it can be considered that the engine load factor KLoff in the reference environment state is corrected based on the representative value representative of the actual engine ambient environment state.

On the other hand, in consideration of the fact that the engine charging efficiency depends on the engine speed NE, in this embodiment, the parameters a ^* (a1 ^* , a2 ^* ), c ^* are set according to the engine speed NE.

Specifically, as shown in FIG. 27A, the gradient a1 ^* increases as the engine speed NE increases when the engine speed NE is low, and the engine speed NE when the engine speed NE is high. It becomes smaller as NE becomes higher. Further, as shown in FIG. 27B, the gradient a2 ^* increases as the engine speed NE increases when the engine speed NE is low, and the engine speed NE increases when the engine speed NE is high. As it gets smaller. These gradients a1 ^* and a2 ^* are obtained in advance by experiments, and are stored in advance in the ROM 34 in the form of maps shown in FIGS. 27A and 27B as functions of the engine speed NE.

Furthermore, as shown in FIG. 28, the engine load factor c ^{* at} the connection point CP increases as the engine speed NE increases when the engine speed NE is low, and the engine speed NE when the engine speed NE is high. Becomes smaller as becomes higher. The engine load factor c ^{* at the} connection point CP is also obtained in advance by experiments, and is stored in advance in the ROM 34 as a function of the engine speed NE in the form of a map shown in FIG.

Therefore, generally speaking, for a plurality of engine speeds NE different from each other, two in-cylinder charged fresh air amount M _c-air at the time of steady operation or intake pipe pressure P _m representing engine load factor KLoff. This means that a linear function formula is obtained and stored in advance.

Then, by detecting the intake pipe pressure P _m by, for example, a pressure sensor 39, equation (26) above from the detected intake pipe pressure P _m or (27) accurately and the engine load factor KLon or KLoff using Thus, the air-fuel ratio can be easily determined, and the air-fuel ratio can be accurately and easily matched with the target air-fuel ratio.

Thus the engine load factor KLon, that the KLoff expressed by a linear function of the intake pipe pressure P _m is the engine load factor KLon, it is not necessary to create a map representing the relationship between the KLoff and the intake pipe pressure P _m Therefore, the effort to create a map is eliminated first. It also means that it is not necessary to solve complicated differential equations, for example, and therefore the calculation load on the CPU 34 is reduced.

By the way, as can be seen by referring to the equation (25), it can be said that the engine load factor KL represents the in-cylinder charged fresh air amount _Mc-air . Here, when the EGR gas is not supplied, considering that only fresh air is filled into the cylinder, the engine load factor KLoff when the EGR gas is not supplied is filled into the cylinder at this time. the total amount of gas, i.e., can be considered to represent the in-cylinder charged gas amount M _c.

Here, assuming that the in-cylinder charged gas amount _Mc does not change between when the EGR gas is supplied and when the EGR gas is not supplied, the engine load factor KLoff when the EGR gas is not supplied is the EGR gas. There not only in-cylinder charged gas amount M _c when being fed, it comes to show respectively a cylinder charged gas amount M _c when the EGR gas is supplied.

On the other hand, as described above, the in _- cylinder charged fresh air amount Mc _-air during steady operation and when EGR gas is supplied is represented by the engine load factor KLon.

Therefore, the result ΔKL (= KLoff−KLon) obtained by subtracting the engine load factor KLon when the EGR gas is supplied from the engine load factor KL when the EGR gas is not supplied is the in-cylinder filling during the steady operation. This means that the EGR gas amount M _c-egr is represented.

Specifically, for example, as shown in FIG. 29, P _m = if a KLoff = KLoff1 when P _m 1 is and KLon = KLon1 is cylinder charged EGR gas during normal operation The quantity M _c-egr is expressed by ΔKL (= KLoff1-KLon1).

Therefore, the in-cylinder charged EGR gas amount M _c-egr during steady operation can be calculated based on the following equation (28).

M _c-egr = kegr1 · ΔKL (28)
Here, kegr1 is a conversion coefficient from the engine load factor KL to the in-cylinder charged EGR gas amount Mc _-egr , and KLoff and KLon are engine load factors calculated from the above equations (26) and (27), respectively. Represents.

Therefore, if the intake pipe pressure P _m is calculated by the above-described calculation formula or is detected by the pressure sensor, the cylinder filling EGR gas at the time of steady operation is calculated from the intake pipe pressure P _m using the above-described formula (28). The quantity M _c-egr can be determined accurately and easily.

Incidentally, as described above, during steady operation, the EGR control valve passage gas flow rate m _egr and the cylinder intake EGR gas amount m _c-egr are equal to each other, and the cylinder filling EGR gas amount M _c-egr is equal to the cylinder intake EGR gas. It is represented by the product of the quantity m _c-egr and ΔT (sec) (M _c-egr = m _c-egr · ΔT).

Therefore, the above-described difference ΔKL also represents the EGR control valve passage gas flow rate m _egr during steady operation.

In the present embodiment, the EGR control valve passage gas flow rate m _egr during steady operation is calculated based on the following equation (29).

m _egr = kegr2 · ΔKL (29)
Here, kegr2 represents a conversion coefficient from the engine load factor KL to the EGR control valve passage gas flow rate m _egr , and KLoff and KLon are engines calculated from the above equations (26) and (27), respectively. It represents the load factor.

As described so far, the EGR control valve passage gas flow rate m _egr at the time of steady operation is calculated using the above equation (29). However, the EGR control valve passage gas flow rate m _egr at the time of transient operation can also be calculated using this equation (29).

That is, the EGR control valve passage gas flow rate m _egr greatly depends on the pressure difference before and after the EGR control valve 22, that is, the difference between the exhaust pressure _Pe and the intake pipe pressure P _m, and is upstream of the EGR control valve 22 during transient operation. given the exhaust pressure P _e and the exhaust temperature T _e is not much different exhaust pressure P _e and the exhaust temperature T _e during steady-state operation, whether transient operation that it would be steady operation, the intake pipe pressure P _m If determined, the EGR control valve passage gas flow rate m _egr is determined.

Therefore, the EGR control valve passage gas flow rate m _egr at the time of steady operation and transient operation can be accurately and easily obtained from the intake pipe pressure P _m using the above-described equation (29). In this case, the in-cylinder charged EGR gas amount M _c-egr at the time of steady operation can be calculated from the EGR control valve passage gas flow rate m _egr at the time of steady operation, or from the difference ΔKL using the above-described equation (28). It can also be calculated.

FIG. 30 shows a routine for calculating the EGR control valve passage gas flow rate m _egr in this embodiment described above. This routine is executed by interruption every predetermined time.

Referring to FIG. 30, first, at step 100, the intake pipe pressure P _m , the engine speed NE, and the EGR opening degree θ _e are read. In the subsequent step 101, the atmospheric temperature correction coefficient ktha and the atmospheric pressure correction coefficient kpa are calculated. In the following step 102, the intake pipe pressure d ^* and the engine load ratios c ^* and r ^* at the connection point CP under the reference environmental state are calculated from the maps of FIG. 23, FIG. 24 (B), and FIG. The In the next step 103, parameters d, c, r are calculated by correcting d ^* , c ^* , r ^* by ktha, kpa. In step 104, the detected intake pipe pressure P _m is whether following the intake pipe pressure d at the connection point is determined. When P _m ≦ d, the routine proceeds to step 105, where a1 ^* and e1 ^* are calculated from the maps of FIG. 22 (C) and FIG. 27 (A). In the subsequent step 106, the gradients a ^* and e ^* are set to a1 ^* and e1 ^* , respectively. Next, the routine proceeds to step 109. On the other hand, when P _m > d, the routine proceeds to step 107, where a2 ^* and e2 ^* are calculated from the maps of FIG. 22 (D) and FIG. 27 (B). In the following step 108, the gradients a ^* and e ^* are set to a2 ^* and e2 ^* , respectively. Next, the routine proceeds to step 109.

In step 109, parameters a and e are calculated by correcting a ^* and e ^* with ktha and kpa. In the subsequent step 110, the engine load factor KLoff is calculated based on the equation (27) (KLoff = a · (P _m −d) + c). In the following step 111, the engine load factor KLon is calculated based on the equation (26) (KLon = e · (P _m −d) + r). In the following step 112, the difference ΔKL is calculated (ΔKL = KLoff−KLon). In the following step 113, the EGR control valve passage gas flow rate m _egr is calculated based on the equation (28) (m _egr = kegr2 · ΔKL).

  In the above-described embodiment, the engine load factors KLoff and KLon are each represented by two linear function equations. However, the engine load factors KLoff and KLon can also be expressed by n m-order function equations (n, m = 1, 2,...), Respectively.

Thus, in the embodiment described above, the first function formula cylinder charging fresh air amount or the engine load factor KLoff when the EGR gas is not being supplied and at the time of steady operation which is a function expression of the intake pipe pressure P _m the first function formula previously obtained is stored a function expression of the cylinder charging fresh air amount or the engine load factor KLon intake pipe pressure P _m when and EGR gas in the time of steady state operation is supplied with representative And the second function equation is obtained and stored in advance, and the in-cylinder filling is performed from the obtained intake pipe pressure P _m by using the first and second function equations. The fresh air amount or engine load factor KLoff, KLon is calculated, the difference ΔKL between the calculated in-cylinder charged fresh air amount or engine load factor KLoff, KLon is calculated, and the EGR control valve passage gas amount m _{egr is set} to the difference ΔKL. Calculation based on And it comes to that.

More generally speaking, the in-cylinder charged fresh air amount or engine load factor KLoff at the time of steady operation and when EGR gas is not supplied, and the in-cylinder at the time of steady operation and when EGR gas is supplied The difference ΔKL from the charged fresh air amount or the engine load factor KLon is expressed by a function expression of the intake pipe pressure P _m , and the function expression is obtained and stored in advance to obtain the intake pipe pressure P _m, and the obtained intake pipe The difference ΔKL is calculated from the pressure P _m using the functional equation, and the EGR control valve passage gas flow rate m _egr during steady operation and transient operation, and the in-cylinder charged EGR gas amount M _c-egr during steady operation, The calculation is based on this difference ΔKL.

  Next, another embodiment of the present invention will be described.

  The above-described difference ΔKL can be expressed by the following equations using equations (27) and (26) representing KLoff and KLon, respectively.

ΔKL = KLoff−KLon
= (A-e) · ( P m -d) + (c-r) (30)
Here, when (a−e) = h and (c−r) = i are replaced, Expression (30) becomes as follows.

ΔKL = h · (P _m −d) + i (31)
h = h1... P _m ≦ d
h = h2... P _m > d

Therefore, as shown in FIG. 31, the difference ΔKL is represented by two linear function expressions of the intake pipe pressure P _m having different gradients and continuous at the connection point CP. That is, the difference ΔKL is expressed by a linear function expression of the gradient h1 when the intake pipe pressure P _m is low, and by a linear function expression of the gradient h2 when the intake pipe pressure P _m is high.

In the present embodiment, two linear function expressions of the intake pipe pressure P _m representing the difference ΔKL are stored in advance in the ROM 34 in the form shown in Expression (31). In this way, the number of parameters can be further reduced.

  Each parameter h, d, i in this equation (31) is calculated based on the following equation.

h1 = h1 ^* · ktha
h2 = h2 ^* · ktha
i = i ^* · ktha · kpa
Here, h1 ^* , h2 ^* , and i ^* are the differences in the gradient and the connection point CP, respectively, when the engine ambient environment state is the reference environment state. These h1 ^*, h2 ^*, i ^* Figure 32 is obtained by experiment in advance, as a function of the engine speed NE and the EGR opening theta _e respectively (A), (B), of a map shown in (C) Is stored in the ROM 34 in advance. The parameter d is the same as that in the above-described embodiment, and thus the description thereof is omitted.

Therefore, generally speaking, for a plurality of different EGR opening degrees θ _e , two linear function expressions of the intake pipe pressure P _m representing the difference ΔKL are respectively obtained and stored in advance. Also, different with respect to a plurality of engine speed NE, it becomes that the two linear functions of the intake pipe pressure P _m representing the difference ΔKL are stored previously obtained one another.

FIG. 33 shows a routine for calculating the EGR control valve passage gas flow rate m _egr in another embodiment of the present invention described above. This routine is executed by interruption every predetermined time.

Referring to FIG. 33, first, at step 120, the intake pipe pressure P _m , the engine speed NE, and the EGR opening degree θ _e are read. In the following step 121, the atmospheric temperature correction coefficient ktha and the atmospheric pressure correction coefficient kpa are calculated. In the subsequent step 122, the intake pipe pressure d ^* and the difference i ^* at the connection point CP under the reference environmental condition are calculated from the maps of FIGS. 23 and 32C. In the following step 123, parameters d and i are calculated by correcting d ^* and i ^* with ktha and kpa. In step 124, the detected intake pipe pressure P _m is whether following the intake pipe pressure d at the connection point is determined. When P _m ≦ d, the routine proceeds to step 125, where h1 ^* is calculated from the map of FIG. In the following step 126, the gradient h ^* is set to h1 ^* . Next, the routine proceeds to step 129. On the other hand, when P _m > d, the routine proceeds to step 127, where h2 ^* is calculated from the map of FIG. In the following step 128, the gradient h ^* is set to h2 ^* . Next, the routine proceeds to step 129.

In step 129, the parameter h is calculated by correcting h ^* with ktha and kpa. In the following step 130, the difference ΔKL is calculated based on the equation (31) (ΔKL = h · (P _m −d) + i). In the following step 131, the EGR control valve passage gas flow rate m _egr is calculated based on the equation (29) (m _egr = kegr2 · ΔKL).

Here, the EGR opening degree θ _e will be briefly described. As described above, the EGR opening degree is represented by the step number STP of the step motor of the EGR control valve 22. That is, when the step number STP becomes zero, the EGR control valve 22 is closed, and when the step number STP increases, the EGR opening. The degree also increases.

However, in practice, as shown in FIG. 34, even when the step number STP increases from zero, the EGR control valve 22 does not open immediately, and when the step number STP exceeds STP1, the EGR control valve 22 is finally opened. I speak. For this reason, it is necessary to represent the EGR opening degree θ _e by the result of subtracting STP1 from the step number STP (STP−STP1).

Further, the EGR control valve 22 usually because it contains manufacturing errors, there is a possibility that actual EGR opening theta _e for the number of steps STP is deviated from the normal opening degree. Therefore, in the internal combustion engine shown in FIG. 1, a correction coefficient kg for making the actual EGR opening coincide with the normal opening is obtained, and this correction coefficient kg is added to the step number STP.

Therefore, the EGR opening degree θ _e is expressed based on the following equation.

θ _e = STP−STP0 + kg
Here, STP0 is the number of steps at which the EGR control valve 22 starts to open in the drawing tolerance center product. In the present embodiment, the EGR opening degree θ _e calculated in this way is used as an argument of the map.

Incidentally, the in-cylinder charged EGR gas amount M _c-egr at the calculated the EGR control valve passage gas flow rate m _egr or during steady-state operation as described above may be further corrected in consideration of the exhaust temperature T _e.

The case of correcting the EGR control valve passage gas flow rate m _egr will be described. In this case, the EGR control valve passage gas flow rate m _egr is expressed by the following equation, for example.

m _egr = m _egr・ kwu ・ krtd ・ kinc
Here, kwu represents a warm-up correction coefficient, krtd represents a retard correction coefficient, and kinc represents an increase correction coefficient.

The warm-up correction coefficient kwu is for correcting the EGR control valve passage gas flow rate m _egr during engine warm-up operation. That is, during the warm-up operation, the exhaust gas temperature _Te is lower than after the completion of the warm-up operation, and the EGR control valve passage gas flow rate m _egr (g / sec) increases accordingly. Since the EGR control valve passage gas flow rate m _egr calculated using the above formula (26), (27) or equation (31) is the EGR control valve passage gas flow rate after the warm-up operation is completed, it is necessary to correct this. There is.

  As shown in FIG. 35 (A), the warm-up correction coefficient kwu decreases as the engine coolant temperature THW representing the degree of warming increases, and becomes 1.0 when the temperature TWU represents the completion of warming up. Retained. The warm-up correction coefficient kwu is stored in advance in the ROM 34 in the form of a map shown in FIG.

On the other hand, the retard correction coefficient krtd is used to correct the EGR control valve passage gas flow rate m _egr when the ignition timing is retarded. That is, the exhaust gas temperature _Te is higher at the time of retard correction than at the time when the retard correction is not performed, and the EGR control valve passage gas flow rate m _egr decreases accordingly.

  As shown in FIG. 35B, the retardation correction coefficient krtd is 1.0 when the retardation amount RTD is zero, and decreases as the retardation amount RTD increases. The retardation correction coefficient krtd is stored in advance in the ROM 34 in the form of a map shown in FIG.

Further, the increase correction coefficient kinc is used to correct the EGR control valve passage gas flow rate m _egr when the fuel injection amount is increased. That is, the exhaust gas temperature _Te is lower during the increase correction than when the increase correction is not performed, and the EGR control valve passage gas flow rate m _egr increases accordingly.

  As shown in FIG. 35C, the increase correction coefficient kinc is 1.0 when the increase correction Finc is zero, and increases as the increase correction Finc increases. This increase correction coefficient kinc is stored in advance in the ROM 34 in the form of a map shown in FIG.

In this way, the EGR control valve passage gas flow rate m _egr can be obtained with higher accuracy.

Incidentally, the engine operating state the exhaust temperature T _e when the ignition timing retard correction or fuel increasing correction is not performed (e.g., the engine speed NE and required load L) determined in advance as a function of the actual exhaust temperature T detecting or estimating _e, so as to correct the EGR control valve passage gas flow rate m _egr based on the difference between the actual exhaust temperature T _e and the exhaust temperature T _e when the retard correction or fuel increasing correction is not performed May be. The same is true for the in-cylinder charged EGR gas amount M _c-egr during the steady operation, and a description thereof is omitted.

  In the internal combustion engine shown in FIG. 1, as described above, the EGR passage 21 downstream of the EGR control valve 22 is branched and connected to the intake branch pipe 12 of each cylinder. In this configuration, in order to suppress variation in the amount of EGR gas supplied to each cylinder, as shown in FIG. 36, a throttle 23 is provided in each EGR passage 21 downstream of the EGR control valve 22. Can do.

In this case, first, during steady operation, the throttle passage gas flow rate mchk (g / sec), which is the flow rate of the EGR gas passing through the throttle 23, matches the EGR control valve passage gas flow rate m _egr . Therefore, as can be seen from the above description, the throttle passage gas flow rate mchk during steady operation can be calculated based on the difference ΔKL. The throttle passage gas flow rate mchk represents the flow rate of the EGR gas flowing into the intake pipe portion.

On the other hand, during transient operation, the throttle passage gas flow rate mchk does not necessarily match the EGR control valve passage gas flow rate m _egr . However, when the volume in the EGR passage 21 from the EGR control valve 22 to the throttle 23 is relatively small, mchk substantially matches m _egr . Accordingly, when the volume in the EGR passage 21 from the EGR control valve 22 to the throttle 23 is relatively small, the throttle passage gas flow rate mchk is calculated based on the difference ΔKL, whether in steady operation or transient operation. It will be possible.

In the present embodiment described so far, for example, the engine load factor KLon, KLoff or the difference ΔKL from the intake pipe pressure P _m estimated using the calculation model or the intake pipe pressure P _m detected by the pressure sensor 39. Is calculated. However, for example, the intake pipe pressure P _m is estimated based on the throttle opening or the output of an air flow meter arranged in the intake duct 14 upstream of the throttle valve 17, and the engine load factor is calculated from the estimated intake pipe pressure P _m. KL can also be calculated.

Here, when the intake pipe pressure P _m is estimated based on the throttle opening, the intake pipe pressure P _m is a function of the throttle opening θ _t , the engine speed NE, and the EGR opening θ _e. It is obtained in advance and stored in the form of a map.

On the other hand, when the intake pipe pressure P _m is estimated based on the output of the air flow meter, the estimated intake pipe pressure P _m is the maximum pressure P that can be taken by the intake pipe pressure P _m for the detection accuracy of the air flow meter. _May exceed _mmax . However, in the region where P _m > P _mmax , as shown in FIG. 37 (B), the engine load factor KLon expressed by the above equation (26) is more than the engine load factor KLoff expressed by the equation (27). In this case, the difference ΔKL becomes a negative value. In other words, the estimated intake pipe pressure P _m is exceeds the maximum pressure P _mmax, there may not be accurately calculating the difference [Delta] KL.

Therefore, as shown in FIG. _37A , in the region where P _m > P _mmax , such a problem can be eliminated if the difference ΔKL is held at a constant value ΔKLC. That is, when the estimated intake pipe pressure P _m exceeds the maximum pressure P _mmax, it is possible to continue to accurately calculate the difference [Delta] KL.

  By the way, when the EGR control valve 22 is closed, the EGR passage 21 downstream of the EGR control valve 22 is filled with air (fresh air). Therefore, immediately after the EGR control valve 22 is opened, fresh air flows into the intake pipe portion from the EGR passage 21 instead of exhaust gas, and thereafter, a mixed gas of fresh air and exhaust gas flows. For a while, this mixed gas flows from the EGR passage 21 into the intake pipe portion.

  Since the temperatures of fresh air and exhaust gas are different, the flow of fresh air from the EGR passage 21 into the intake pipe portion is considered to particularly affect the relationship established in terms of energy conservation.

  However, in the above-described embodiment, it is not considered that fresh air may flow into the intake pipe portion from the EGR passage 21 as described above.

  In view of this, in the above-described embodiment, in consideration of this, a relational expression that is established in terms of energy conservation between the air and exhaust gas flowing into the intake pipe portion and the air and exhaust gas flowing out from the intake pipe portion into the cylinder. As follows, the following equation (32) may be adopted.

In equation (32), when the EGR control valve passage gas flow rate m _egr is the flow rate of the gas flowing into the intake pipe portion from the EGR passage 21, m _egr-egr flows into the intake pipe portion from the EGR passage 21. The flow rate of exhaust gas in gas (flow rate of inflow exhaust gas), and m _egr-air is the flow rate of air (fresh air) out of the gas flowing from the EGR passage 21 into the intake pipe portion (hereinafter referred to as inflow). This is referred to as a fresh air flow.

In Equation (32), m _c-egr is the flow rate of exhaust gas flowing into the cylinder (cylinder inflow exhaust gas flow rate), and m _c-air is the flow rate of air (fresh air) flowing into the cylinder. (Hereinafter referred to as in-cylinder incoming fresh air flow rate).

Here, a method of calculating the inflow exhaust gas partial flow rate m _egr-egr and the inflow fresh _air partial flow rate m _egr-air will be described.

When the EGR control valve 22 is opened, the flow rate of the gas flowing from the EGR passage 21 into the intake pipe portion is equal to the EGR control valve passing gas flow rate. As described above, the inflowing exhaust gas partial flow rate m _egr-egr starts to increase after a certain period of time (dead time) has elapsed, and then gradually increases with a first-order delay. Therefore, the inflow fresh _air flow rate m _egr-air is obtained from the following equation (33).

Here, the EGR control valve passage gas flow rate m _egr is obtained as described in the above-described embodiment. Therefore, if the flow rate m _{egr-egr for} the inflowing exhaust gas is obtained, the inflowing fresh _air flow rate m _egr-air is also obtained.

As described above, the inflowing exhaust gas partial flow rate m _egr-egr is obtained by calculating the dead time Td1 and the time constant τ1 from the functions shown in FIGS. 16A and 16B stored in the ROM 34 in the form of a map. It is calculated as a change starting toward the EGR control valve passage gas flow rate after the dead time Td1 elapses, and thereafter changing toward the EGR control valve passage gas flow rate with a time constant τ1.

On the other hand, the in-cylinder inflow exhaust gas flow rate _mc-egr and the in-cylinder inflow fresh air flow rate _{mc -air} are obtained in the same manner.

That is, the in-cylinder inflow exhaust gas flow rate _mc-egr starts to increase after a certain period of time (dead time) has elapsed, and then gradually increases with a primary delay. Therefore, the in-cylinder inflow exhaust gas flow rate _mc-air is obtained from the following equation (34).

As can be seen from this equation, if the in-cylinder inflow exhaust gas flow rate mc _-egr is obtained, the in-cylinder inflow fresh air flow rate _mc-air is also obtained.

As described above, the cylinder inflow exhaust gas flow rate _mc-egr is determined from the functions shown in FIGS. 17A and 17B stored in the ROM 34 in the form of a map, and the dead time Td2 and the time constant τ2. And after the dead time Td2 elapses, the change starts toward the EGR control valve passage gas flow rate, and thereafter, the change is calculated as changing toward the EGR control valve passage gas flow rate with the time constant τ2. .

  Thus, according to this embodiment, since the fresh air flowing into the intake pipe portion from the EGR passage 21 is also taken into account, the accuracy of the calculated value is improved.

  By the way, when the distance from the part where the EGR passage 21 is connected to the intake pipe part to the intake valve is much shorter than the distance from the EGR control valve 22 to the part, the following expression (35) is established.

That is, the flow rate m _c-egr of the exhaust gas flowing into the cylinder is substantially equal to the flow rate m _egr-egr of the exhaust gas (for the inflowing exhaust gas) flowing into the intake pipe portion from the EGR passage 21.

  Therefore, in this case, the above equation (32) can be rewritten as the following equation (36).

  Therefore, in the above-described embodiment, the above equation (36) is established as a relational expression that is established on the energy conservation law between the air and exhaust gas flowing into the intake pipe portion and the air and exhaust gas flowing out from the intake pipe portion into the cylinder. ) May be adopted.

Here, the flow rate of fresh air flowing into the intake pipe portion from the EGR passage 21 (flow rate of fresh air flow) m _egr-air is obtained from the above equation (33) if the flow rate of incoming exhaust gas m _egr-egr is found. .

Further, the flow rate of fresh air flowing into the cylinder (cylinder inflow fresh air flow rate) _mc-air can be obtained from the above equation (34) if the in _- cylinder inflow exhaust gas flow rate mc _-egr is obtained.

Of course, when the inflowing exhaust gas partial flow rate m _egr-egr and the in-cylinder inflowing exhaust gas flow rate _mc-egr are calculated when the EGR control valve passage gas flow rate m _egr changes, as described above, it is a waste. Time and first order lag are considered.

Thus, as a parameter to be input to a formula (36), it does not include the exhaust temperature T _e, in the case of employing the equation (36) does not need to determine the exhaust temperature T _e.

  By the way, the above equation (36) can be rewritten as the following equation (37).

  Here, from the relationship of the above equation (33), the above equation (37) can be further rewritten as the following equation (38).

  Further, the equation (38) can be rewritten as the following equation (39).

Here, when the engine operating state is in a steady state, because the intake pipe temperature the T _m approximately equal to the atmospheric temperature _{T a (T m ≒ T a} ), the last term in the right side of the equation (39)

Becomes zero.

On the other hand, if the engine operating state is in the transient state is different from the intake pipe temperature The T _m is the atmospheric temperature T _a, the greater the change in the intake pipe negative pressure P _m, the difference between these temperatures is increased. However, if the EGR control valve passage gas flow rate m _egr is large, at this time, the intake pipe pressure P _m is kept high to promote combustion in the cylinder, and conversely, the EGR control valve passage gas flow rate m _egr is even less, at this time, even if the difference between the intake pipe temperature T _m and the atmospheric temperature T _a is large, the effect of which is smaller.

  Therefore, even when the engine operating state is a transient state, the last term on the right side of Equation (39)

Can be approximated to zero.

  Summing up the above, the equation (39) can be approximated as the following equation (40).

  Therefore, in the above-described embodiment, the above equation (40) is established as a relational expression that is established in the energy conservation law between the air and exhaust gas flowing into the intake pipe portion and the air and exhaust gas flowing out from the intake pipe portion into the cylinder. ) May be adopted.

  According to this, it is not necessary to consider the inflow exhaust gas partial flow rate, the cylinder inflow exhaust gas flow rate, the inflow fresh air flow rate, and the cylinder inflow fresh air flow rate.

  By the way, the flow rate of exhaust gas flowing into the intake pipe portion (flow rate of inflowing exhaust gas) is equal to the flow rate of gas passing through the EGR control valve, and the flow rate of exhaust gas flowing into the cylinder (flow rate of exhaust gas in the cylinder) flows in. If it is equal to the exhaust gas partial flow rate, the above equation (15) is rewritten to the following equation (41), and the above equation (16) is rewritten to the following equation (42).

Here, T _k is the temperature of the exhaust gas or air flowing into the intake pipe portion from the EGR passage 21 or a mixed gas thereof.

  Furthermore, Expression (42) can be rewritten as the following Expression (43).

  Therefore, the above-described condition, that is, the condition that the inflow exhaust gas partial flow rate is equal to the EGR control valve passage gas flow rate and the in-cylinder inflow exhaust gas flow rate is equal to the inflow exhaust gas partial flow rate is satisfied, or this condition Is allowed as a relational expression that holds in the law of conservation of mass and the law of conservation of energy between the air and exhaust gas flowing into the intake pipe portion and the air and exhaust gas flowing out from the intake pipe portion to the cylinder, The above formulas (41) and (43) may be adopted.

According to this, it is not necessary to obtain the EGR control valve passage gas flow rate m _egr .

In the equations (41) and (43), the in-cylinder inflow fresh air flow rate _mc-air is calculated from the engine load factor KL obtained from the above equation (26) or (27). Specifically, the engine load factor KL is calculated from the above equation (26) or (27), the engine load factor KL is input to the equation (25), and by solving this equation, the cylinder inflow fresh air amount M _c-air is calculated, finally, by dividing the time required for one cycle of the cylinder inflow fresh air amount M _c-air to the internal combustion engine at the time divided by the number of cylinders, the cylinder inflow fresh air flow rate m _{c -air} is calculated.

Incidentally, in the above embodiment, the throttle valve passage air flow rate m _t is calculated by using the formula of equation (1). In this equation (1), the function Φ (P _m / P _a ) is used. Considering that the shape of the throttle valve 18 is complicated or pulsation occurs in the air that has passed through the throttle valve 18. reduces the accuracy of calculation of the function [Phi, therefore, likely also falls calculation accuracy of this function the throttle valve passing air flow rate is calculated by using the [Phi m _t.

Therefore, as these functions Φ (P _m / P _a) the throttle valve passage air flow rate calculation method can be eliminated calculation error of adopting a method of calculating the throttle valve passage air flow rate m _t using the following equation (44) You can also.

Here, P _m-egr is the intake pipe pressure that will eventually converge, and is determined based on the engine speed NE, the throttle opening θ _t , and the EGR opening θ _e (P _m-egr = F1 (NE, θ _t , θ _e )).

  Further, e, d, and r are adaptation parameters, and details of these adaptation parameters will be described later.

Φ (P _m / P _a ) is a function of the equation (2), and Φ (P _m-egr / P _a ) is _a variable P _m / P _a in this function Φ (P _m / P _a ). it is a function instead of the _m-egr / P _a.

  Next, the process of deriving the above equation (44) will be described.

When representing the throttle valve passage air flow rate during steady-state operation in m _t-TA, equation for the throttle valve passage air flow rate m _t-TA can be expressed as the following equation (45).

  Then, by dividing each side of equation (1) by each side of equation (45) and arranging it, the following equation (46) is derived.

On the other hand, as described above, the in-cylinder inflow air flow rate _mc-air is proportional to the in _- cylinder inflow air amount _Mc-air , and the in-cylinder inflow air amount _Mc-air is proportional to the engine load factor KL. Since the engine load factor KL is obtained from the above equation (26), the cylinder inflow air flow rate _mc-air at the time of steady operation and when the EGR gas is supplied is expressed by the following equation (47). M _c-air = f · (P _m −d) + g (47)
f = f1... P _m ≦ d
f = f2... P _m > d

Here, f is a conforming parameter (f = f ₂ (NE, θ _e )) determined based on the engine speed NE and the EGR opening θ _e, and d is a conforming parameter determined based on the engine speed NE ( d = f ₃ (NE)), and g is a fitting parameter (g = f ₄ (NE, θ _e )) determined based on the engine speed NE and the EGR opening θ _e . In other words, it can be said that f, d, and g are coefficients with parameters representing the engine operating state as variables.

Here, in the steady operation, the in-cylinder inflow air flow rate _mc-air is equal to the throttle valve passing air flow rate mt _-TA, and the intake pipe pressure P _m is finally converged in the transient operation. Since it becomes equal to the pipe pressure P _m-egr , the following equation (48) is derived from the equation (47).

m _t−TA = f · (P _m−egr −d) + g (48)
f = f1... P _m-egr ≦ d
f = f2 _Pm-egr > d

  Then, by substituting the equation (48) into the equation (46), the equation (44) is derived.

The intake pipe pressure P _m-egr is a function of the engine speed NE, the throttle opening θ _t , and the EGR opening θ _e (P _m-egr = f ₁ (NE, θ _t , θ _e )). These are obtained in advance by experiments or the like and stored in the ROM 34 in the form of a map.

Further, the adaptation parameter f is obtained in advance by experiments or the like as a function of the engine speed NE and the EGR opening degree θ _e (f = f ₂ (NE, θ _e )) and is stored in the ROM 34 in the form of a map. .

The adaptation parameter d is obtained in advance by experiments or the like as a function of the engine speed NE (d = f ₃ (NE)) and is stored in the ROM 34 in the form of a map.

The conforming parameter g is obtained in advance by experiments or the like as a function of the engine speed NE and the EGR opening degree θ _e (g = f ₄ (NE, θ _e )) and is stored in the ROM 34 in the form of a map. .

Therefore, in the present embodiment, the intake pipe pressure that will eventually converge during the transient operation based on the engine speed NE, the throttle opening degree θ _t , the EGR opening degree θ _{e, and the} like from each map described above. P _m-egr and conforming parameters f, d, f are calculated. On the other hand, the intake pipe pressure P _m is calculated based on the detected value of the intake pipe pressure sensor 40, and these are input to the above equation (44). by solving the equation, the throttle valve passage air flow rate m _t is calculated.

The intake pipe pressure P _m is estimated or calculated based on other means, for example, the throttle opening θ _t , or in the case where an air flow meter is provided, estimated based on the detected value of the air flow meter Or may be calculated.

According to the present embodiment, in the steady operation, the intake pipe pressure P _m and the intake pipe pressure P _m-egr are equal (P _m = P _m-egr ), so in the equation (44), Φ (P _m / P _a ) / Φ (P _m−egr / P _a ) becomes 1, and the function Φ disappears, so even if the calculation accuracy of the value of the function Φ is low, at least during steady operation the accuracy of the throttle valve passage air flow rate m _t calculated by using the equation (44) is good.

Incidentally, the EGR gas is supplied into the cylinder mainly during the engine low load operation. In the engine low load operation, the exhaust pressure P _e is said to be approximately equal to the atmospheric pressure P _a. Therefore, when EGR gas is supplied into the cylinder, the pressure upstream of the EGR control valve 22 is equal to the pressure upstream of the throttle valve 18. Since the pressure downstream of the EGR control valve 22 and the pressure downstream of the throttle valve 18 are the intake pipe pressure P _m, they are equal in the first place.

  That is, when the EGR gas is supplied into the cylinder, the pressure upstream and downstream of the EGR control valve 22 is equal to the pressure upstream and downstream of the throttle valve 18, respectively. In this case, the EGR opening can be converted into the throttle opening equivalent to it relatively easily. Next, the conversion method of this EGR opening degree is demonstrated.

That is, if the pressures upstream and downstream of the EGR control valve 22 are equal to the pressures upstream and downstream of the throttle valve 18 and the engine speed NE is constant, several different EGR opening degrees θ _e (= 0, θ _e 1, θ The relationship between the throttle opening θ _t and the intake pipe pressure P _{m-egr in e} 2, θ _e 3) is as shown in FIG.

That is, if the EGR opening degree is the same, the more the throttle opening is large, the negative pressure P _m-egr intake pipe increases toward the atmospheric pressure P _a. If the throttle opening is the same, the intake pipe negative pressure P _m-egr increases toward the atmospheric pressure as the EGR opening increases.

Since the intake pipe pressure P _m-egr tends to decrease as the engine speed NE increases, in the example shown in FIG. 38, the throttle opening at each EGR opening increases as the engine speed NE increases. The curve indicating the relationship with the intake pipe pressure moves downward as a whole.

Referring to FIG. 38, it can be seen that when the throttle opening is zero and the EGR opening is θ _e 1, the intake pipe pressure is a value P _m-egr 1. On the other hand, although the intake pipe pressure is this value P _m-egr 1, when the EGR opening is zero (θ _e = 0), it can be seen that the throttle opening is θ _t 1.

That is, when the intake pipe pressure is the same value P _m-egr 1, if the throttle opening is zero, the EGR opening is θ _e 1, and conversely, if the EGR opening is zero, the throttle The opening is θ _t 1.

Therefore, in the example shown in FIG. 38, when the EGR opening degree θ _e 1 is converted into the throttle opening degree, the value is θ _t 1. Similarly, in the example shown in FIG. 38, when the EGR opening θ _e 2 is converted into the throttle opening, the value θ _t 2 is obtained, and when the EGR opening θ 3 is converted into the throttle opening, the value θ _t 3 is obtained. .

  The relationship between such an EGR opening and an equivalent throttle opening (equivalent throttle opening) is as shown in FIG. 39. Here, the equivalent throttle opening is almost a linear function of the EGR opening. ing. That is, if the relationship as shown in FIG. 39 is used, the throttle opening equivalent to the EGR opening can be obtained relatively easily.

  Therefore, the equivalent throttle opening is obtained in advance in the form of a map shown in FIG. 39 by experiments or the like and stored in the ROM 34, and the above-described conditions, that is, the pressures upstream and downstream of the EGR control valve 22 are respectively controlled by the throttle valve When the condition equal to the upstream / downstream pressure is satisfied, or when this condition is allowed, the equivalent throttle opening is calculated from this map using the EGR opening as an argument, and this equivalent throttle opening is calculated as the actual throttle opening. The total throttle opening may be used by adding the degree.

  According to this, in the above-described embodiment, it is not necessary to use the EGR opening as a parameter, so that the map creation effort and the calculation load of the map or the calculation formula are reduced.

  Of course, since the EGR opening is not directly used as a parameter, for example, the map argument is reduced by one. In general, when the map argument is decreased by one, a value calculated from the map is obtained. However, according to the above-described method, since the EGR opening is substantially included in the total throttle opening, the accuracy of the value calculated from the map or the calculation formula is high. .

  As described above, when the engine speed NE increases, the curve indicating the relationship between the throttle opening and the intake pipe pressure at each EGR opening only moves downward, so that the intake pipe The relationship between the EGR opening and the throttle opening with the same pressure does not change much.

  Therefore, the function as shown in FIG. 39 is substantially the same even if the engine speed NE fluctuates. Therefore, according to the above-described method, the EGR opening can be converted into the equivalent throttle opening only by using one map as shown in FIG.

  Of course, in order to improve the conversion accuracy to the equivalent throttle opening, a correction coefficient for the calculated equivalent throttle opening is obtained using the engine speed NE as a variable, and, for example, by multiplying the equivalent throttle opening by the correction coefficient. The influence of the engine speed on the equivalent throttle opening may be compensated.

  Further, the relationship shown in FIG. 39 is a relationship derived with reference to a position where the throttle opening is zero. For example, the relationship shown in FIG. The relationship shown in FIG. 39 is derived.

Further, when this method is adopted, for example, the intake pipe pressure P _m-egr is previously tested as a function of the engine speed NE and the throttle opening θ _t (P _m-egr = f ₅ (NE, θ _t )). Although sought by like are stored in the ROM34 in the form of a map, in an internal combustion engine this map having no EGR device, the intake pipe at the time of steady state operation from the engine speed NE and the throttle opening theta _t pressure P _{m- Since} the map is similar to the map for calculating _TA , the map for calculating the intake pipe pressure P _m-TA can be used as the map for calculating the intake pipe pressure P _m-egr .

Meanwhile, between the opening cross-sectional area A _t of the throttle opening theta _t and the throttle valve 18, non-linear relationship as shown in FIG. 40.

  Here, when considering a method of converting the above-described EGR opening into an equivalent throttle opening, in this method, the actual throttle opening, that is, the original throttle opening before the equivalent throttle opening is added. Regardless of the value, the equivalent throttle opening is calculated from the EGR opening using the relationship shown in FIG. The equivalent throttle opening is added to the actual throttle opening.

However, as seen by reference to FIG. 40, as the throttle opening theta _t is changed by the same, by the original throttle opening theta _t, the change amount of the opening cross-sectional area A _t of the throttle valve 18 is different. Rate of change of the opening cross-sectional area A _t of the throttle valve per unit throttle opening is in the region throttle opening is small, a relatively small, in the region throttle opening is large, a relatively large.

  That is, even if the same equivalent throttle opening is added to the actual throttle opening, the degree of influence varies depending on the actual throttle opening, that is, the original throttle opening.

  On the other hand, the slope Sb at the throttle opening (in the above example, zero or the throttle opening during idling operation) used as a reference when creating the map shown in FIG. 39 is obtained on the curve of FIG. When a value Sb / S obtained by dividing the slope Sb by the slope S at each throttle opening obtained from the curve of FIG. 40 is plotted as a function of the throttle opening, a curve shown by the solid line of FIG. 41 is obtained. .

A value Sb / S obtained by dividing the slope Sb by the slope S is a correction coefficient for the equivalent throttle opening with respect to the actual throttle opening. That is, for example, when the actual throttle opening is θ _t 1, the equivalent throttle opening is multiplied by a correction coefficient K1 corresponding to the throttle opening θ _t 1 in FIG. of the equivalent throttle opening is added to the actual throttle opening, by calculating the total throttle opening, the throttle opening degree of the total, with the opening cross-sectional area a _t of the throttle opening theta _t and the throttle valve 18 The opening degree takes into consideration the non-linear relationship between them.

According to this, when the equivalent throttle opening is calculated, since a non-linear relationship between the opening cross-sectional area A _t of the throttle opening theta _t and the throttle valve 18 is considered, the equivalent throttle opening The accuracy of the value calculated using the total throttle opening to which is added is improved.

Incidentally, as shown in Figure 41, the throttle opening theta _t is very large area, even if the throttle opening theta _t increases, the throttle opening area A _t is no longer becoming more large, the gradient S is zero Therefore, in this region, the correction coefficient Sb / S becomes infinite.

On the other hand, in this region, it is known that the intake pipe pressure P _m-egr hardly changes even if the throttle opening θ _t changes or the EGR opening θ _e changes. There is no problem even if the equivalent throttle opening is considered to be very small or almost zero.

Therefore, in the very large region throttle opening theta _t, as shown by the broken line in FIG. 41, the correction coefficient K may be gradually reduced values toward zero. This avoids inconveniences in control, maintains continuity of control, and matches the actual phenomenon.

It is a figure which shows the whole internal combustion engine provided with the control apparatus of this invention. It is a figure which shows the cylinder inflow gas amount model applicable to the internal combustion engine which is not equipped with an EGR apparatus. It is a figure which shows the relationship between a throttle opening and a flow coefficient. It is a figure which shows the relationship between a throttle opening and opening cross-sectional area. Is a diagram showing the function Φ a (P _m / P _a). It is a figure which shows the basic concept of a throttle model. It is a figure which shows the basic concept of the intake pipe model applicable to the internal combustion engine which is not equipped with an EGR apparatus. It is a figure which shows the basic concept of an intake valve model. It is a figure regarding the definition of the cylinder filling gas amount and the cylinder inflow gas flow rate. It is a figure which shows the cylinder inflow gas amount model of this invention applicable to an internal combustion engine provided with the EGR apparatus. It is a figure which shows the basic concept of the intake pipe model applicable to the internal combustion engine provided with the EGR apparatus. It is a figure which shows the basic concept in calculation of an EGR control valve passage gas flow rate. It is a figure showing the relationship between an engine load factor and exhaust pressure. It is a figure showing the relationship between an EGR control valve passage gas flow rate and exhaust temperature. It is a figure which shows the change of an inflow exhaust gas partial flow rate when an EGR control valve passage gas flow rate changes. It is a figure which shows the dead time Td1 regarding the inflow exhaust gas partial flow rate, and time constant (tau) 1. It is a figure which shows the dead time Td2 and time constant (tau) 2 regarding cylinder inflow exhaust gas flow volume. It is a figure for _{demonstrating the} EGR control valve passage gas amount m _egr . Exhaust pressure P _e, is a graph showing the exhaust temperature T _e, and P _e / √T _e. It is a diagram which shows a function ( _Pm / _Pe ). Is a diagram showing an example of a relationship between the engine load factor KLon and the intake pipe pressure P _m. It is a diagram which shows gradient e1, e2. It is a diagram which shows the intake pipe pressure d in a connection point. It is a diagram which shows the engine load factor r in a connection point. Is a diagram showing an example of a relationship between the engine load factor KLon and the intake pipe pressure P _m. Is a diagram showing an example of a relationship between the engine load ratio KLoff and the intake pipe pressure P _m. It is a diagram which shows gradient a1, a2. It is a diagram which shows the engine load factor c in a connection point. It is a diagram for demonstrating difference (DELTA) KL. It is a flowchart which shows the calculation routine of the EGR control valve passage gas flow rate m _egr . Is a diagram showing an example of the relationship between the difference ΔKL the intake pipe pressure P _m. It is a diagram which shows the difference i in gradient h1, h2 and a connection point. It is a flowchart which shows the calculation routine of the EGR control valve passage gas flow rate m _egr by another embodiment of this invention. It is a diagram which shows the relationship between an EGR opening degree and step number (theta) _e . It is a diagram which shows various correction coefficients. FIG. 6 is a partial view of an internal combustion engine showing still another embodiment of the present invention. Intake pipe pressure P _m is a diagram showing the engine load factor KL when exceeding the maximum pressure P _mmax. It is a figure which shows the relationship between throttle opening (theta) _t and intake pipe pressure _Pm-egr . It is a diagram showing the relationship between the EGR opening degree theta _e and its equivalent of the throttle opening theta _t. Is a diagram showing the relationship between the opening cross-sectional area A _t of the throttle opening theta _t and the throttle valve. It is a figure which shows the relationship between actual throttle opening (theta) _t and the correction coefficient K. FIG.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Engine body 5 Combustion chamber 6 Intake valve 7 Intake port 8 Exhaust valve 11 Fuel injection valve 13 Intake pipe 18 Throttle valve 22 EGR control valve

Claims (3)

  An exhaust gas flow control valve for controlling the flow rate of the exhaust gas flowing in the exhaust gas recirculation passage and connecting the exhaust passage to the intake passage downstream of the throttle valve via the exhaust gas recirculation passage is provided in the exhaust gas recirculation passage. In the arranged internal combustion engine, an air amount calculating means for repeatedly calculating the amount of air passing through the throttle valve, an exhaust gas amount calculating means for repeatedly calculating the amount of exhaust gas passing through the exhaust gas flow rate control valve, A cylinder gas amount calculating means for repeatedly calculating the amount of gas flowing into the cylinder, and repeatedly calculating the pressure in the intake passage downstream of the throttle valve based on at least the air amount, the exhaust gas amount, and the cylinder gas amount And the exhaust gas amount calculating means calculates a next exhaust gas amount based on at least the previous pressure in the intake passage. The amount of exhaust gas passing through the exhaust gas flow control valve is in the cylinder when the flow rate of exhaust gas flowing into the cylinder is zero during steady operation at the pressure in the intake passage downstream of the same throttle valve. The pressure calculation device is calculated from the difference between the flow rate of air flowing into the cylinder and the flow rate of air flowing into the cylinder when the flow rate of exhaust gas flowing into the cylinder is not zero during steady operation.   2. The pressure according to claim 1, wherein the difference is calculated by two linear function expressions of pressures in the intake passage downstream of the throttle valve, the gradients being different from each other and continuing at the connection point. Calculation device. The pressure calculation device according to claim 1, wherein the temperature of the exhaust gas that passes through the exhaust gas flow control valve and flows into the intake passage downstream of the throttle valve is calculated from a map having a parameter representing an engine operating state as a variable .

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