JP2008051005A - Control device for internal combustion engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a control device for an internal combustion engine capable of controlling ignition timing to provide maximum net torque. <P>SOLUTION: This control device for the internal combustion engine is provided with a means calculating ignition timing maximizing indicated torque, a means calculating friction torque FMEP based on an engine operation condition, and a means correcting ignition timing maximizing indicated torque calculated by the means to maximize net torque provided by subtracting friction torque from the indicated torque. <P>COPYRIGHT: (C)2008,JPO&INPIT

Description

  The present invention relates to an internal combustion engine control apparatus, and more particularly to an apparatus capable of suitably controlling an ignition timing of an internal combustion engine.

  Generally, in a control device for an internal combustion engine, each state quantity representing an engine operating state such as a rotation speed and a load is detected, and on the basis of each state quantity, by referring to a map previously adapted through experiments or the like, The ignition timing is determined.

  However, in recent engines, various variable mechanisms are provided, and there are many components, and variations in the combustion state are likely to occur. Therefore, the number of operating states of the engine has become enormous, and it is no longer difficult to set the ignition timing in advance so as to be compatible with all these operating states.

  Therefore, in recent years, it has been considered to determine an optimum ignition timing for obtaining the maximum torque, that is, MBT (Minimum Advance for Best Torque) for each model, and to perform feedback control so that the actual ignition timing matches this MBT. (For example, refer to Patent Document 1). According to this, it is possible to always bring the ignition timing close to the optimal ignition timing without being affected by the solid variation, and also to create various maps, that is, adaptation work, which has conventionally spent a lot of time and labor. There is an advantage that the development period can be significantly shortened. In this MBT control, it is indispensable to acquire in-cylinder information. For this reason, a sensor capable of detecting in-cylinder information such as an in-cylinder pressure sensor is provided.

JP 2005-146888 A

  However, the conventional optimum ignition timing, that is, MBT, is an ignition timing at which the indicated torque based on the in-cylinder pressure and the in-cylinder volume is maximized. In other words, the ignition timing is such that the maximum torque can be obtained for the combustion itself occurring in the combustion chamber.

  On the other hand, the performance required for the engine is that the torque actually output from the crankshaft, that is, the net torque is maximized. Thus, in view of this point of view, the ignition timing at which the net torque is maximized was examined, and it was found that the ignition timing does not necessarily match the ignition timing at which the indicated torque is maximized.

  Therefore, the present invention has been made in view of such circumstances, and an object thereof is to provide a control device for an internal combustion engine capable of controlling the ignition timing so that the maximum net torque can be obtained.

  In order to achieve the above object, an internal combustion engine control apparatus according to an aspect of the present invention includes a maximum indicated torque ignition timing calculating means for calculating an ignition timing that maximizes the indicated torque, and a friction based on the engine operating state. Friction torque calculating means for calculating torque, and ignition timing correcting means for correcting the ignition timing calculated by the maximum indicated torque ignition timing calculating means so that the net torque obtained by subtracting the friction torque from the indicated torque is maximized. It is characterized by comprising.

  Preferably, an indicated torque calculating means for calculating indicated torque based on an engine operating state is further provided, and the ignition timing correcting means is calculated by the friction torque calculated by the friction torque calculating means and the indicated torque calculating means. An ignition timing correction amount is calculated based on the indicated torque.

  Preferably, the ignition timing correction means calculates the ignition timing correction amount based on a ratio between the friction torque and the indicated torque.

  Preferably, an in-cylinder pressure detecting means for detecting an in-cylinder pressure and a rotational speed detecting means for detecting an engine rotational speed are further provided, and the friction torque calculating means is configured to detect the maximum in-cylinder pressure detected by the in-cylinder pressure detecting means. The friction torque is calculated based on the value and the engine rotational speed detected by the rotational speed detecting means.

  Preferably, the maximum indicated torque ignition timing calculating means calculates an ignition timing at which a combustion ratio at a predetermined crank angle becomes a predetermined value.

  According to the present invention, it is possible to provide a control device for an internal combustion engine capable of controlling the ignition timing so that the maximum net torque can be obtained.

  The best mode for carrying out the present invention will be described below with reference to the accompanying drawings.

  FIG. 1 is a schematic diagram showing the configuration of the present embodiment. As shown in the figure, the internal combustion engine 1 generates power by burning a mixture of fuel and air inside a combustion chamber 3 formed in a cylinder block 2 and reciprocating a piston 4 in the combustion chamber 3. To do. The internal combustion engine 1 is a vehicular multi-cylinder engine (only one cylinder is shown), and is a spark ignition type internal combustion engine, more specifically, a gasoline engine.

  In the cylinder head of the internal combustion engine 1, an intake valve Vi for opening and closing the intake port and an exhaust valve Ve for opening and closing the exhaust port are provided for each cylinder. Each intake valve Vi and each exhaust valve Ve are opened and closed by a camshaft (not shown). Here, the intake valve Vi is provided with a variable valve timing mechanism (hereinafter referred to as “VVTi”) 21 so that the opening / closing timing of the intake valve Vi can be changed. A spark plug 7 for igniting the air-fuel mixture in the combustion chamber 3 is attached to the top of the cylinder head for each cylinder. Further, an injector (fuel injection valve) 12 is disposed in the cylinder head for each cylinder so that fuel is directly injected into the combustion chamber 3. The piston 4 is configured as a so-called deep dish top surface type, and a concave portion 4a is formed on the upper surface thereof. In the internal combustion engine 1, fuel is directly injected from the injector 12 toward the concave portion 4 a of the piston 4 in a state where air is sucked into the combustion chamber 3. As a result, a layer of a mixture of fuel and air is formed (stratified) in the vicinity of the spark plug 7 and separated from the surrounding air layer, and stable stratified combustion is executed.

  The intake port of each cylinder is connected to a surge tank 8 serving as an intake air collecting chamber via a branch pipe for each cylinder. An intake pipe 13 that forms an intake manifold passage is connected to the upstream side of the surge tank 8, and an air cleaner 9 is provided at the upstream end of the intake pipe 13. An air flow meter 5 for detecting the intake air amount and an electronically controlled throttle valve 10 are incorporated in the intake pipe 13 in order from the upstream side. An intake passage is formed by the intake port, the surge tank 8 and the intake pipe 13.

  On the other hand, the exhaust port of each cylinder is connected to an exhaust pipe 6 forming an exhaust collecting passage through a branch pipe for each cylinder, and a catalyst 11 made of a three-way catalyst is attached to the exhaust pipe 6. An exhaust passage is formed by the exhaust port, the branch pipe, and the exhaust pipe 6. A pre-catalyst sensor 17 for detecting the exhaust air / fuel ratio is installed upstream of the catalyst 11.

  The spark plug 7, throttle valve 10, injector 12, VVTi 21, and the like described above are electrically connected to an electronic control unit (hereinafter referred to as “ECU”) 20 as control means. The ECU 20 includes a CPU, a ROM, a RAM, an input / output port, a storage device, and the like, all not shown. In addition to the air flow meter 5 and the pre-catalyst sensor 17, the ECU 20 includes a crank angle sensor 14 that detects the crank angle (phase) of the internal combustion engine 1 and an accelerator opening that detects the accelerator opening, as shown in the figure. A sensor 15, an intake pressure sensor 16 for detecting the intake pressure, a throttle opening sensor 19 for detecting the opening of the throttle valve 10, and other various sensors are electrically connected via an A / D converter (not shown) or the like. Yes. The ECU 20 controls the ignition plug 7, the throttle valve 10, the injector 12, the VVTi 21, etc. so as to obtain a desired output based on the detection values of various sensors, etc., and the ignition timing, fuel injection amount, fuel injection timing, Controls throttle opening, intake valve opening / closing timing, etc. The throttle opening is normally controlled to an opening corresponding to the accelerator opening, and the intake valve opening / closing timing is controlled to a timing according to the engine operating state.

  Each cylinder is provided with an in-cylinder pressure sensor 22 including a semiconductor element, a piezoelectric element, an optical fiber detection element, or the like. Each in-cylinder pressure sensor 22 is disposed on the cylinder head so that the pressure receiving surface faces the corresponding combustion chamber 3, and is electrically connected to the ECU 20 via an A / D converter (not shown). Each in-cylinder pressure sensor 22 gives a voltage signal proportional to the in-cylinder pressure of the corresponding combustion chamber 3 to the ECU 20. The detection value of each in-cylinder pressure sensor 22 is sequentially given to the ECU 20 every minute time, and is updated and stored in a predetermined storage area (buffer) of the ECU 20 by a predetermined amount.

  Next, ignition timing control in this embodiment will be described.

In the present embodiment, the basic ignition timing is controlled to the optimal ignition timing θ _ZMBT that is predetermined for each engine model and that maximizes the indicated torque. This ignition timing θ _ZMBT is the conventional MBT and is referred to as “illustrated torque MBT” for convenience. Specifically, the basic ignition timing is controlled so that the combustion ratio MFB at a predetermined reference crank angle θ0 becomes a predetermined value φ. In the present embodiment, φ = 50% and θ0 = 8 ° ATDC are set. However, the crank angle θ0 indicating MFB = φ when the indicated torque is maximum is adapted according to the engine model. The combustion ratio MFB is the ratio of the amount of heat generated at that time to the total amount of heat generated per one cylinder cycle (taken through each stroke of intake, compression, expansion, and exhaust). The combustion rate MFB is calculated by the ECU 20 every minute time as described later.

  Here, the illustrated torque will be described. FIG. 2 is a PV diagram of the internal combustion engine, wherein the vertical axis indicates the cylinder pressure Pc, the horizontal axis indicates the cylinder volume V, and Vtdc and Vbdc indicate the cylinder volumes at the top dead center and the bottom dead center, respectively. As is known, the diagram changes as indicated by an arrow. The area indicated by A corresponds to the amount of work performed by the engine, and the area indicated by B corresponds to the pumping loss. Here, the indicated torque Tz is defined as Tz = A−B. In order to obtain a large indicated torque Tz, it is necessary to increase the area of the region A and reduce the area of the region B.

  By the way, the ignition timing that realizes MFB = φ when the crank angle is θ0 as described above is the optimal ignition timing in the sense of maximizing the indicated torque Tz here, in other words, the combustion in the combustion chamber This is the optimum ignition timing when only the above is considered. However, in an actual engine, a value obtained by subtracting a mechanical loss such as friction such as friction torque from the indicated torque generated in the combustion chamber, that is, a net torque is taken out as an output from the crankshaft, and actually the vehicle The net torque is used for driving and the like. Therefore, it is preferable to control the ignition timing to an ignition timing at which the net torque is maximized (hereinafter referred to as “net torque MBT”) rather than an ignition timing at which the indicated torque is maximized (illustrated torque MBT). .

When the indicated torque MBT and the net torque MBT were examined, it was found that they did not necessarily match. FIG. 3 shows the relationship between the indicated torque and the net torque for each ignition timing. As shown in the figure, the net torque MBTθ _SMBT and the _indicated torque MBTθ _ZMBT are at different times, and there is an error between them. The friction torque, which is the difference between the indicated torque and the net torque, changes according to the ignition timing, and the friction torque increases as the _indicated torque MBTθZMBT _advances . The indicated torque MBTθ _ZMBT becomes a value on the advance side, and the friction torque increases when, for example, the engine load is low, the valve overlap is large, or the engine temperature is low. It is. In the region where the indicated torque MBTθ _ZMBT is a value on the advance side, the influence of the error between the indicated torque MBTθ _ZMBT and the net torque MBTθ _SMBT cannot be ignored, and this error also causes a decrease in engine efficiency.

Therefore, in the present embodiment, the ignition timing is controlled to coincide with the net torque MBTθ _SMBT by correcting the indicated torque MBT as follows.

  FIG. 4 shows a routine for calculating the ignition timing in the ignition timing control of the present embodiment. This calculation process is executed by the ECU 20 for each minute crank angle (for example, every 10 °). Each timing in this calculation process is shown in FIG. FIG. 5 shows a change in the in-cylinder pressure Pc with respect to a crank angle θ in a predetermined one cylinder and one cycle of the engine. As shown in the drawing, ignition is performed before a predetermined angle of compression top dead center TDC (θ = 0 °), and the air-fuel mixture in the combustion chamber is combusted by this ignition, and the in-cylinder pressure Pc is increased after compression top dead center. The maximum pressure Pcmax is reached at a predetermined angle, and then decreases.

  Referring to FIG. 4, first, in step S101, it is determined whether or not the exhaust angle TDC (θ = −360 °, timing indicated by I in FIG. 5) is detected by the crank angle sensor 14. If not detected, step S101 is executed again to wait until it is detected, and if detected, the process proceeds to step S102, and the valve timings of the intake valve Vi and the exhaust valve Ve are acquired. Here, since the ECU 20 controls the VVTi 21 based on the engine operating state (for example, the rotation speed and the load), the ECU 20 acquires the current opening / closing timing of the intake valve Vi based on the position of the VVTi21. Further, the ECU 20 acquires the opening / closing timing of the exhaust valve Ve stored in advance as a constant value.

  Next, in steps S103 to S105, the crank angle acquisition and in-cylinder pressure Pc detection for each minute crank angle are repeated until the crank angle θ reaches the intake valve closing timing IVC (the timing indicated by II in FIG. 5). That is, in step S103, the value of the current crank angle θ detected by the crank angle sensor 14 is acquired, and in step S104, the value of the in-cylinder pressure Pc detected by the in-cylinder pressure sensor 22 at the time of the crank angle θ. To be acquired. The acquired value of the in-cylinder pressure Pc is stored in the buffer in association with the value of the crank angle θ. In step S105, it is determined whether or not the crank angle θ has reached the intake valve closing timing IVC. If the crank angle θ has not reached the intake valve closing timing IVC, steps S103 and S104 are repeatedly executed, and the crank angle θ If the intake valve closing timing IVC has been reached, the process proceeds to step S106.

  Next, in steps S106 to S110, the crank angle θ and the in-cylinder pressure Pc are repeatedly detected and stored until the crank angle θ reaches the exhaust valve opening timing EVO (the timing indicated by III in FIG. 5). For each crank angle, the current heating rate dQ and the cumulative heating rate Q, which is the cumulative value, are calculated. That is, in step S106, the current value of the crank angle θ detected by the crank angle sensor 14 is acquired, and in step S107, the value of the in-cylinder pressure Pc detected by the in-cylinder pressure sensor 22 at the time of the crank angle θ. To be acquired. The acquired value of the in-cylinder pressure Pc is stored in the buffer in association with the value of the crank angle θ.

  In step S108, the current heating rate dQ is calculated based on the following equation (1).

κ is the specific heat ratio of the in-cylinder mixture, and is treated as a constant value here. For example, κ = 1.32. V is an in-cylinder volume and is a function of the crank angle θ, and is calculated by the ECU 20 based on the current crank angle θ.

  In the next step S109, the cumulative heating rate Q that is the cumulative value or integral value of the heating rate dQ up to the present time is calculated. That is, the currently calculated heating rate dQ is added to the previously calculated cumulative value of the heating rate dQ to calculate the current cumulative heating rate Q. The value of the cumulative heating rate Q calculated so far is stored in the buffer in association with the value of the crank angle θ at the current time.

  In step S110, it is determined whether or not the crank angle θ has reached the exhaust valve opening timing EVO. If the crank angle θ has not reached the exhaust valve opening timing EVO, steps S106 to S109 are repeatedly executed to determine the crank angle θ. If the exhaust valve opening timing EVO is reached, the process proceeds to step S111. Thus, the calculation of the cumulative heating rate Q is repeatedly performed from the intake valve closing timing IVC to the exhaust valve opening timing EVO. The maximum cumulative heating rate Qmax when the crank angle θ reaches the exhaust valve opening timing EVO is expressed by the following equation (2).

Next, in step S111, a crank angle θ ₅₀ at which the combustion ratio MFB = Q / Qmax, which is the ratio of the cumulative heating rate Q and the maximum cumulative heating rate Qmax, becomes φ (= 50%) is calculated. That is, the combustion ratio MFB for each crank angle from the intake valve closing timing IVC to the exhaust valve opening timing EVO is calculated based on the stored value of the buffer, and the calculated combustion ratio MFB is accumulated with φ as φ. A crank angle corresponding to the heating rate Q is calculated as θ ₅₀ .

Here, FIG. 6 shows how the combustion ratio MFB = Q / Qmax changes with respect to the crank angle θ. As shown in the figure, as the crank angle θ increases, that is, as the air-fuel mixture in the cylinder ignites and combustion proceeds, the combustion rate MFB gradually increases from zero and finally converges to 1. In step S111, a crank angle θ ₅₀ corresponding to MFB = φ = 0.5 is calculated.

  Next, in steps S112 to S114, until the crank angle θ reaches the exhaust bottom dead center (θ = 180 °, timing indicated by IV in FIG. 5), crank angle acquisition and in-cylinder pressure Pc detection for each minute crank angle are performed. Repeatedly executed. That is, in step S112, the current value of the crank angle θ detected by the crank angle sensor 14 is acquired. In step S113, the value of the in-cylinder pressure Pc detected by the in-cylinder pressure sensor 22 at the time of the crank angle θ. To be acquired. The acquired value of the in-cylinder pressure Pc is stored in the buffer in association with the value of the crank angle θ. In step S114, it is determined whether or not the crank angle θ has reached the exhaust bottom dead center (exhaust BDC). If the crank angle θ has not reached the exhaust bottom dead center, steps S112 and S113 are repeatedly executed. If the angle θ has reached exhaust bottom dead center, the process proceeds to step S115.

  In step S115, the indicated torque IMEP of the cycle is calculated based on the following equation (3).

That is, here, the product value Pc · dV / dθ of the in-cylinder pressure and the in-cylinder volume at each minute crank angle is the exhaust bottom dead center from the intake bottom dead center INBDC (θ = −180 °, timing indicated by V in FIG. 5). The indicated torque IMEP is calculated by integration up to the point EXBDC (θ = 180 °, timing indicated by IV in FIG. 5).

  As described with reference to FIG. 2, the indicated torque in a strict sense is represented by Tz = A−B, which is a value obtained by subtracting the area of the region B from the area of the region A. Therefore, in order to accurately obtain the indicated torque Tz, such integration must be performed over the entire cycle. However, the pumping loss (B) may be considered constant, for example, when the engine is in steady operation. Here, for the sake of simplicity, only the region A corresponding to the engine work is focused, and the indicated torque IMEP is defined by the area of the region from the intake bottom dead center INBDC to the exhaust bottom dead center EXBDC. However, even if the integration process is executed over one cycle, there is no problem in that the pumping loss (B) is reduced from the engine work (A) as a constant value or a value corresponding to the engine operating state. May be.

  Next, in step S116, the value of the maximum in-cylinder pressure Pcmax in the cycle is acquired based on the stored in-cylinder pressure data, and the value of the engine speed NE calculated based on the output of the crank angle sensor 14 is acquired. Is done. Based on the maximum in-cylinder pressure Pcmax and the rotational speed NE, the friction torque FMEP is calculated according to the following equation (4).

C _{1 to} C ₄ are constants adapted in advance. Since the torque has a linear relationship with the average effective pressure, and the two are equivalent, the torque may be replaced with the average effective pressure. Here, the average effective pressure is a value obtained by dividing work by the stroke volume.

Next, in step S118, based on the values of the indicated torque IMEP calculated in step S115 and the friction torque FMEP calculated in step S117, this is a correction amount for matching the indicated torque MBTθ _ZMBT with the net torque MBTθ _SMBT. An ignition timing correction amount ΔSA is calculated. In this calculation, first, a torque ratio R _MEP that is a ratio of the indicated torque IMEP and the friction torque FMEP is calculated by the following equation (5).

Thereafter, the ignition timing correction amount ΔSA is calculated based on the calculated torque ratio R _MEP according to a predetermined correction amount map shown in FIG. According to this map, a larger correction amount ΔSA is obtained in a region where the torque ratio R _MEP is smaller than a region where the torque ratio R _MEP is large, and a larger correction amount ΔSA is obtained as the torque ratio R _MEP is smaller. What this means is that, as the friction torque FMEP is relatively large with respect to indicated torque IMEP, is that the net torque MBTθ _SMBT is delayed from the indicated torque MBTθ _ZMBT. For example, when the engine operating state is high rotation or low load, the friction torque FMEP is relatively increased, and the degree of correction is increased. The map may be replaced with a function.

When the ignition timing correction amount ΔSA is calculated in this way, next, in step S119, the ignition timing SA _next of the next cycle is calculated by the following equation (6).

SA is the ignition timing of the current cycle. Specifically, it is the ignition timing calculated in step S119 of the previous cycle and actually ignited in the current cycle. It should be noted here that the values SA _next , SA, and ΔSA indicating the ignition timing and the correction amount thereof are BTDC based on the compression top dead center TDC, while the values θ ₅₀ , θ0 indicating the crank angle are used. The point is that the unit is ATDC. These relationships are shown in FIG.

Thus, the processing of the current cycle is finished, and the ignition timing SA _next of the next cycle that is finally calculated becomes the ignition timing at which ignition is actually executed in the next cycle.

Here, the ignition timing SA _next of the next cycle represented by Equation 6 will be described with reference to FIG. The part of the right side {SA + (θ ₅₀ −θ0)} in Equation ₆ is a feedback part for making the ignition timing coincide with the _indicated torque MBTθ _ZMBT , as in the conventional case. Further, the portion of the right side {−ΔSA} of Equation 6 is a portion for correcting the _indicated torque MBTθ _ZMBT so that the final ignition timing is matched with the net torque MBTθ _SMBT .

In FIG. 8, the solid line shows the change in the combustion ratio MFB = Q / Qmax with respect to the crank angle θ when ignition is performed with the _indicated torque MBTθ _ZMBT . As can be seen, in this case, the combustion ratio MFB is φ (0.5 = 50%) at the reference crank angle θ0 (8 ° ATDC).

On the other hand, it is assumed that the change in the combustion ratio MFB as a result of ignition at the ignition timing SA of the current cycle is as indicated by a one-dot chain line. In this case, the crank angle θ ₅₀ at the time when the combustion ratio MFB becomes φ is shifted at a time earlier than the reference crank angle θ0. The _indicated torque MBTθ _ZMBT in the next cycle can be obtained by adding the amount of deviation (θ ₅₀ −θ0) of the ignition timing to the ignition timing SA of the current cycle. In the illustrated example, since (θ ₅₀ −θ0) is a negative value, the finally obtained value is a value smaller than the ignition timing SA of the current cycle, that is, a value on the retard side closer to the compression top dead center TDC. Thus, the deviation amount of the ignition timing of the current cycle with respect to the indicated torque MBT is fed back to the next cycle.

Further, in the present embodiment, the ignition timing {SA + (θ ₅₀ −θ0)} that maximizes the indicated torque is further corrected in consideration of the friction torque. That is, the final ignition timing SA _next is calculated by subtracting the correction amount ΔSA from the ignition timing {SA + (θ ₅₀ −θ0)} that maximizes the indicated torque. As described above, as the friction torque FMEP increases relative to the indicated torque IMEP, the correction amount ΔSA increases, and the finally obtained ignition timing SA _next is made closer to the compression top dead center TDC or retarded. It is changed to the value on the side. For example, as can be seen from Equation 4 above, the higher the engine speed NE, the greater the friction torque FMEP, so the ignition timing is corrected to the retard side. Further, the lower the engine load (or the smaller the indicated torque), the greater the friction torque FMEP relative to the indicated torque IMEP, so that the ignition timing is corrected to the retard side.

Thus, according to the present embodiment, the ignition timing at which the indicated torque becomes maximum is corrected so that the net torque becomes maximum, in other words, the ignition timing is controlled so as to coincide with the net torque MBTθ _SMBT. Is done. Therefore, the maximum torque can be actually obtained from the crankshaft, the engine efficiency can be increased, and the improvement of fuel consumption can be achieved.

In the calculation of the ignition timing SA _next of the next cycle in step S119, correction based on valve timing or valve overlap may be added.

  In calculating the heating rate dQ in step S108, the specific heat ratio κ calculated according to the following equation 7 may be used.

Here, GA is an intake air amount detected by the air flow meter 5 or calculated based on an air model, AF is an air-fuel ratio and a value calculated from AF = GA / GF (where GF is a fuel injection amount), κ _air is the specific heat ratio of air, κ _fuel is the specific heat ratio of the fuel, κ _res is the specific heat ratio of the residual gas (mixed gas of CO ₂ , H ₂ O, and N ₂ ) in the cylinder, R is the gas constant, and T is the in-cylinder gas temperature, Fd is the mass percentage of the residual gas in the cylinder interior gas (burned gas cycle before remaining in the cylinder) is, c ₁₁ to c ₃₆ is a predetermined coefficient. P ₁ and V ₁ are the in-cylinder pressure and the in-cylinder volume at the heating rate dQ calculation timing in step S108 (that is, current), and P ₂ and V ₂ are the in-cylinder pressure at the calculation timing one time before the calculation timing. And the in-cylinder volume. M is a value obtained according to the following equation (8). Such a specific heat ratio κ can be calculated using an approximate expression based on data in the JANAF table.

  In the embodiment, the ECU 20 constitutes the maximum indicated torque ignition timing calculating means, the friction torque calculating means, the ignition timing correcting means, and the indicated torque calculating means. In-cylinder pressure sensor 22 constitutes in-cylinder pressure detection means, and crank angle sensor 14 and ECU 20 constitute rotation speed detection means.

  Although a preferred embodiment of the present invention has been described above, various other embodiments of the present invention are possible. For example, in addition to the in-cylinder pressure sensor 22, an ion sensor can be used as means for detecting the in-cylinder combustion state. Combustion in the cylinder is a chemical reaction, and ions are generated during combustion. Therefore, the combustion state in the cylinder can be detected by detecting the amount of generated ions with an ion sensor.

  The embodiment of the present invention is not limited to the above-described embodiment, and includes all modifications, applications, and equivalents included in the concept of the present invention defined by the claims. Therefore, the present invention should not be construed as being limited, and can be applied to any other technique belonging to the scope of the idea of the present invention.

It is the schematic which shows the structure of this embodiment. It is a PV diagram of an internal combustion engine. It is a graph which shows the relationship between illustrated torque, net torque, friction torque, and ignition timing. It is a flowchart which shows the calculation process routine of the ignition timing in the ignition timing control of this embodiment. It is a cycle diagram for showing each timing in the calculation processing of FIG. It is a graph which shows the relationship between combustion ratio MFB = Q / Qmax and crank angle (theta). An ignition timing correction amount map is shown. It is a graph for demonstrating the calculation method of the ignition timing of the next cycle.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Internal combustion engine 3 Combustion chamber 7 Spark plug 12 Injector 14 Crank angle sensor 20 Electronic control unit (ECU)
22 In-cylinder pressure sensor θ _{ZMBT Indicated} torque MBT
θ _SMBT net torque MBT
ΔSA Ignition timing correction amount FMEP Friction torque IMEP Indicated torque R _MEP torque ratio Pc In-cylinder pressure Pcmax Maximum in-cylinder pressure θ Crank angle V In-cylinder volume MFB Combustion rate φ Predetermined value of combustion rate θ0 Reference crank angle

Claims (5)

A maximum indicated torque ignition timing calculating means for calculating an ignition timing such that the indicated torque is maximized;
Friction torque calculating means for calculating friction torque based on the engine operating state;
An ignition timing correction means for correcting the ignition timing calculated by the maximum indicated torque ignition timing calculation means so that the net torque obtained by subtracting the friction torque from the indicated torque is maximized. Engine control device. Further comprising an indicated torque calculating means for calculating the indicated torque based on the engine operating state;
The ignition timing correction means calculates an ignition timing correction amount based on the friction torque calculated by the friction torque calculation means and the indicated torque calculated by the indicated torque calculation means. The internal combustion engine control device described. The control apparatus for an internal combustion engine according to claim 2, wherein the ignition timing correction means calculates the ignition timing correction amount based on a ratio between the friction torque and the indicated torque. In-cylinder pressure detecting means for detecting in-cylinder pressure, and rotation speed detecting means for detecting engine rotation speed,
The friction torque calculating means calculates the friction torque based on the maximum value of the in-cylinder pressure detected by the in-cylinder pressure detecting means and the engine rotational speed detected by the rotational speed detecting means. The control device for an internal combustion engine according to any one of claims 1 to 3. The control device for an internal combustion engine according to any one of claims 1 to 4, wherein the maximum indicated torque ignition timing calculation means calculates an ignition timing such that a combustion ratio at a predetermined crank angle becomes a predetermined value.

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