Patent Publication Number: US-7216626-B2

Title: Control apparatus for internal combustion engine

Description:
This nonprovisional application is based on Japanese Patent Applications Nos. 2004-328143 and 2005-078461 filed with the Japan Patent Office on Nov. 11, 2004 and Mar. 18, 2005, respectively, the entire contents of which are hereby incorporated by reference. 
   BACKGROUND OF THE INVENTION 
   1. Field of the Invention 
   The present invention relates to a control apparatus for an internal combustion engine having a first fuel injection mechanism (an in-cylinder injector) for injecting a fuel into a cylinder and a second fuel injection mechanism (an intake manifold injector) for injecting a fuel into an intake manifold or an intake port, and relates particularly to a technique for determining a timing of ignition with a fuel injection ratio between the first and second fuel injection mechanisms considered. 
   2. Description of the Background Art 
   An internal combustion engine having an intake manifold injector for injecting a fuel into an intake manifold of the engine and an in-cylinder injector for always injecting a fuel into a combustion chamber of the engine, and configured to stop fuel injection from the intake manifold injector when the engine load is lower than a preset load and to cause fuel injection from the intake manifold injector when the engine load is higher than the set load, is known. 
   In such an internal combustion engine, one configured to switch between stratified charge combustion and homogeneous combustion in accordance with its operation state is known. In the stratified charge combustion, the fuel is injected from the in-cylinder injector during a compression stroke to form a stratified air-fuel mixture locally around a spark plug, for lean combustion of the fuel. In the homogeneous combustion, the fuel is diffused in the combustion chamber to form a homogeneous air-fuel mixture, for combustion of the fuel. 
   Japanese Patent Laying-Open No. 2001-020837 discloses a fuel injection control apparatus for an engine that switches between stratified charge combustion and homogeneous combustion in accordance with an operation state and that has a main fuel injection valve for injecting a fuel directly into a combustion chamber and a secondary fuel injection valve for injecting a fuel into an intake port of each cylinder. This fuel injection control apparatus for the engine is characterized in that the fuel injection ratio between the main fuel injection valve and the secondary fuel injection valve is set in a variable manner based on an operation state of the engine. 
   According to this fuel injection control apparatus for the engine, the stratified charge combustion is carried out using only the main fuel injection valve directly injecting the fuel into the combustion chamber, while the homogeneous combustion is carried out using both the main fuel injection valve and the secondary fuel injection valve (or using only the secondary fuel injection valve in some cases). This can keep the capacity of the main fuel injection valve small, even in the case of an engine of high power. Linearity in injection duration/injection quantity characteristic of the main fuel injection valve in a low-load region such as during idling is improved, which in turn improves accuracy in control of the fuel injection quantity. Accordingly, it is possible to maintain favorable stratified charge combustion, and thus to improve stability of the low-load operation such as idling. In the homogeneous combustion, both the main and secondary fuel injection valves are employed, so that the benefit of the direct fuel injection and the benefit of the intake port injection are both enjoyed. Accordingly, favorable homogeneous combustion can also be maintained. 
   In the fuel injection control apparatus for the engine disclosed in Japanese Patent Laying-Open No. 2001-020837, the stratified charge combustion and the homogeneous combustion are employed according-to the situations, which complicates ignition control, injection control and throttle control, and requires control programs corresponding to the respective combustion manners. Particularly, upon switching between the combustion manners, these controls require considerable changes, making it difficult to realize desirable controls (of fuel efficiency, emission purification performance) at the time of transition. Further, in the stratified combustion region where lean combustion is carried out, the three-way catalyst does not work, in which case a lean NOx catalyst needs to be used, leading to an increased cost. 
   Based on the foregoing, an engine has also been developed which does not provide stratified charge combustion, and thus does not need control for switching between the stratified charge combustion and the homogeneous combustion and does not require an expensive lean NOx catalyst. 
   In controlling the engine to be ignited with its coolant having lower temperature, spark advance is introduced for correction. This is because when the coolant has lower temperature (poorer atomization is provided) lower combustion rates are provided and the engine is less prone to knock. The spark advance can provide an increased period of time between ignition and exhaust, and despite lower combustion rates the air fuel mixture can sufficiently be combusted. 
   In a cold state, however, for a range shared by an intake manifold injector injecting fuel into an intake manifold (or port) having low temperature in a cold state and an in-cylinder injector injecting the fuel into a cylinder having high temperature despite the cold state to both inject the fuel the fuel is atomized differently. As such, the fuel injected through the in-cylinder injector and that through the intake manifold injector have different conditions, respectively. The fuel atomized differently forms air fuel mixtures having different conditions and hence combusting differently, and the air fuel mixtures combusting differently require different optimal timings of ignition. As such, using the coolant&#39;s temperature alone to calculate an amount of spark advance cannot provide an accurate timing of ignition (or an accurate amount of spark advance). Furthermore, not only in the cold state but a transitional period from the cold state to a warm state as well, for a range shared by the intake manifold and in-cylinder injectors to both inject the fuel the temperature in the cylinder and that at the intake port increase at different rates. As such, using the coolant&#39;s temperature alone to calculate an amount of spark advance cannot provide an accurate amount of spark advance. Note that Japanese Patent Laying-open No. 2001-20837 only discloses that each injector is driven to achieve a fuel injection ratio corresponding to the operation state of interest and a timing of ignition is set, and the document does not provide a solution to the problem described above. 
   SUMMARY OF THE INVENTION 
   An object of the present invention is to provide a control apparatus for an internal combustion engine having first and second fuel injection mechanisms bearing shares, respectively, of injecting fuel into a cylinder and an intake manifold, respectively, that can calculate an accurate amount of variation introduced to time ignition in a cold state and a transitional period from the cold state to a warm state when the fuel injection mechanisms share injecting the fuel. 
   The present invention in one aspect provides a control apparatus for an internal combustion engine that controls an internal combustion engine having a first fuel injection mechanism injecting fuel into a cylinder and a second fuel injection mechanism injecting the fuel into an intake manifold. The control apparatus includes: a controller controlling the first and second fuel injection mechanisms to bear shares, respectively, of injecting the fuel at a ratio calculated as based on a condition required for the internal combustion engine; a detector detecting a temperature of the internal combustion engine; and an ignition timing controller controlling an ignition device to vary a timing of ignition. The ignition timing controller uses the ratio and the temperature to calculate an amount of variation introduced to time the internal combustion engine to be ignited in a cold state and applies the amount to control the ignition device to vary the timing of ignition. 
   In the present invention, for a range shared by the first fuel injection mechanism (e.g., an in-cylinder injector) and the second fuel injection mechanism (e.g., an intake manifold injector) to both inject the fuel the cylinder&#39;s interior and the intake port increase in temperature at different rates. In a cold state and a transitional period from the cold state to a warm state, because of this difference in temperature, spark advance or retard is introduced at different degrees. The ignition timing controller considers a ratio between the fuel injected into the cylinder and that injected into the intake port and calculates as based on the internal combustion engine&#39;s temperature (e.g., that of a coolant of an engine) an amount of spark advance or retard (collectively referred to as an amount of variation introduced to time ignition) in the cold state. Thus the internal combustion engine having two fuel injection mechanisms that share injecting fuel into different portions can have an accurate spark advance or retard in the cold state. Thus a control apparatus for an internal combustion engine can be provided that can calculate an accurate amount of variation introduced to time ignition in a cold state and a transitional period from the cold state to a warm state when fuel injection mechanisms share injecting the fuel. 
   The present invention in another aspect provides a control apparatus for an internal combustion engine that controls an internal combustion engine having a first fuel injection mechanism injecting fuel into a cylinder and a second fuel injection mechanism injecting the fuel into an intake manifold. The control apparatus includes: a controller controlling the first and second fuel injection mechanisms to bear shares, respectively, of injecting the fuel at a ratio calculated as based on a condition required for the internal combustion engine; a detector detecting a temperature of the internal combustion engine; a calculator calculating a reference timing of ignition; and an ignition timing controller using an amount of variation introduced to time ignition to control an ignition device to vary the reference timing of ignition. The ignition timing controller uses the ratio and the temperature to calculate an amount of variation introduced to time the internal combustion engine to be ignited in a cold state and applies the amount to control the ignition device to vary the reference timing of ignition. 
   In the present invention for a range shared by the first fuel injection mechanism (e.g., an in-cylinder injector) and the second fuel injection mechanism (e.g., an intake manifold injector) to both inject the fuel the cylinder&#39;s interior and the intake port increase in temperature at different rates. In a cold state and a transitional period from the cold state to a warm state, because of this difference in temperature, spark advance or retard is introduced at different degrees. The ignition timing controller considers a ratio between the fuel injected into the cylinder and that injected into the intake port and calculates as based on the internal combustion engine&#39;s temperature (e.g., that of a coolant of an engine) an amount of spark advance or retard for correction in the cold state. This amount is used to vary a reference timing of ignition calculated as based on the internal combustion engine&#39;s operation state. Thus the internal combustion engine having two fuel injection mechanisms that share injecting fuel into different portions can achieve an accurately varied timing of ignition in the cold state. Thus a control apparatus for an internal combustion engine can be provided that can calculate an accurate amount of variation introduced to time ignition in a cold state and a transitional period from the cold state to a warm state when fuel injection mechanisms share injecting the fuel. 
   The present invention in still another aspect provides a control apparatus for an internal combustion engine that controls an internal combustion engine having a first fuel injection mechanism injecting fuel into a cylinder and a second fuel injection mechanism injecting the fuel into an intake manifold. The control apparatus includes: a controller controlling the first and second fuel injection mechanisms to bear shares, respectively, of injecting the fuel at a ratio calculated as based on a condition required for the internal combustion engine; a detector detecting a temperature of the internal combustion engine; and an ignition timing controller controlling an ignition device to vary a timing of ignition. The ignition timing controller uses the ratio and the temperature to calculate an amount of spark advance of the internal combustion engine in a cold state and applies the amount to control the ignition device to vary the timing of ignition. 
   In the present invention, for a range shared by the first fuel injection mechanism (e.g., an in-cylinder injector) and the second fuel injection mechanism (e.g., an intake manifold injector) to both inject the fuel the cylinder&#39;s interior and the intake port increase in temperature at different rates. In a cold state and a transitional period from the cold state to a warm state, because of this difference in temperature, spark advance is introduced at different degrees. The ignition timing controller considers a ratio between the fuel injected into the cylinder and that injected into the intake port and calculates as based on the internal combustion engine&#39;s temperature (e.g., that of a coolant of an engine) an amount of spark advance in the cold state. Thus the internal combustion engine having two fuel injection mechanisms that share injecting fuel into different portions can have an accurate spark advance in the cold state. Thus a control apparatus for an internal combustion engine can be provided that can calculate an accurate amount of spark advance in a cold state and a transitional period from the cold state to a warm state when fuel injection mechanisms share injecting the fuel. 
   The present invention in still another aspect provides a control apparatus for an internal combustion engine that controls an internal combustion engine having a first fuel injection mechanism injecting fuel into a cylinder and a second fuel injection mechanism injecting the fuel into an intake manifold. The control apparatus includes: a controller controlling the first and second fuel injection mechanisms to bear shares, respectively, of injecting the fuel at a ratio calculated as based on a condition required for the internal combustion engine; a detector detecting a temperature of the internal combustion engine; a calculator calculating a reference timing of ignition; and an ignition timing controller using an amount of spark advance for correction to control an ignition device to vary the reference timing of ignition. The ignition timing controller uses the ratio and the temperature to calculate an amount of spark advance of the internal combustion engine for correction in a cold state and applies the amount to control the ignition device to vary the reference timing of ignition. 
   In the present invention, for a range shared by the first fuel injection mechanism (e.g., an in-cylinder injector) and the second fuel injection mechanism (e.g., an intake manifold injector) to both inject the fuel the cylinder&#39;s interior and the intake port increase in temperature at different rates. In a cold state and a transitional period from the cold state to a warm state, because of this difference in temperature, spark advance is introduced at different degrees. The ignition timing controller considers a ratio between the fuel injected into the cylinder and that injected into the intake port and calculates as based on the internal combustion engine&#39;s temperature (e.g., that of a coolant of an engine) an amount of spark advance for correction in the cold state. This amount is used to vary a reference timing of ignition calculated as based on the internal combustion engine&#39;s operation state. Thus the internal combustion engine having two fuel injection mechanisms that share injecting fuel into different portions can have an accurate spark advance in the cold state. Thus a control apparatus for an internal combustion engine can be provided that can calculate an accurate amount of spark advance in a cold state and a transitional period from the cold state to a warm state when fuel injection mechanisms share injecting the fuel. 
   Preferably the ignition timing controller calculates the amount of spark advance to be decreased when the first fuel injection mechanism is increased in the ratio. 
   In accordance with the present invention, as the first fuel injection mechanism an in-cylinder injector injecting fuel into a cylinder exists, and the cylinder&#39;s internal temperature is higher than the intake port&#39;s temperature. As such, if the in-cylinder injector injects the fuel at higher ratios, it is not necessary to introduce a significant spark advance. Despite small spark advance, combustion as desired can be achieved. 
   Still preferably the ignition timing controller calculates the amount of spark advance to be increased when the second fuel injection mechanism is increased in the ratio. 
   In accordance with the present invention, as the second fuel injection mechanism an intake manifold injector injecting fuel into an intake manifold exists, and the intake port&#39;s temperature is lower than the cylinder&#39;s internal temperature. As such, if the intake manifold injector injects the fuel at higher ratios, significant spark advance can be introduced to achieve combustion as desired. 
   Still preferably the ignition timing controller calculates the amount of spark advance to be decreased when the temperature is increased. 
   In accordance with the present invention higher temperatures in the internal combustion engine help the fuel to atomize. As such, a large spark advance is not required and despite a small spark advance combustion as desired can be achieved. 
   Still preferably the ignition timing controller calculates the amount of spark advance to be increased when the temperature is decreased. 
   In accordance with the present invention lower temperatures in the internal combustion engine prevent the fuel from atomizing. Accordingly, a large spark advance is introduced so that combustion as desired can be achieved. 
   The present invention in still another aspect provides a control apparatus for an internal combustion engine that controls an internal combustion engine having a first fuel injection mechanism injecting fuel into a cylinder and a second fuel injection mechanism injecting the fuel into an intake manifold. The control apparatus includes: a controller controlling the first and second fuel injection mechanisms to bear shares, respectively, of injecting the fuel at a ratio calculated as based on a condition required for the internal combustion engine, the ratio including preventing one of the fuel injection mechanisms from injecting the fuel; a detector detecting a temperature of the internal combustion engine; a storage storing a reference timing of ignition and an amount of variation introduced to time ignition in a cold state; and an ignition timing controller controlling an ignition device by varying the reference timing of ignition by the amount to provide a varied timing of ignition. The storage stores the amount of variation introduced to time the internal combustion engine to be ignited in the cold state, the amount being calculated as based on a condition of a mixture of the fuel injected through the fuel injection mechanism and air. 
   In accordance with the present invention, in a cold state, for a range shared by the first fuel injection mechanism (e.g., an in-cylinder injector) and the second fuel injection mechanism (e.g., an intake manifold injector) to both inject the fuel the cylinder receiving the fuel injected through the first fuel injection mechanism and the intake manifold receiving the fuel injected through the second fuel injection mechanism are high and low, respectively, in temperature, which contributes to a difference between how the fuel is atomized in the cylinder (satisfactorily) and how it is atomized in the intake manifold (unsatisfactorily) and hence a difference between a mixture of the fuel in the cylinder and air and that of the fuel in the intake manifold and air. More specifically the difference is for one reason caused by how differently the fuel is atomized in a cold state. The in-cylinder injector injects the fuel into the hot cylinder even in the cold state. As such, the fuel can be atomized satisfactorily and thus mixed with air to form a satisfactorily homogeneous air fuel mixture allowing a high combustion rate. In contrast, the intake manifold injector injects the fuel into the cold intake manifold in the cold state. As such, the fuel is atomized insufficiently and thus mixed with air to form an inhomogeneous air fuel mixture resulting in a low combustion rate. Accordingly the ignition device is controlled to achieve an optimal timing of ignition by varying a reference timing of ignition by an amount of variation introduced to time ignition that is calculated as based on an air fuel mixture&#39;s condition (such as how it is atomized, how homogeneous it is, and the like) that reflects the fuel injection mechanisms&#39; fuel injection ratio and in addition thereto the internal combustion engine&#39;s temperature and the like. Thus the internal combustion engine having two fuel injection mechanisms that share injecting fuel and provide air fuel mixtures having different conditions can achieve an accurately set timing of ignition. As a result a control apparatus that can calculate an accurate timing of ignition can be provided for an internal combustion engine having first and second fuel injection mechanisms bearing shares, respectively, of injecting fuel to inject the fuel into a cylinder and an intake manifold, respectively, that are implemented by two types of fuel injection mechanisms injecting fuel differently. 
   Preferably the storage stores the amount calculated from the temperature and the ratio. 
   In accordance with the present invention the temperature and the ratio contribute to how the fuel is atomized, which in turn contributes to how an air fuel mixture is formed, and therefrom an amount of variation introduced to time ignition is calculated. The ignition device is controlled by varying a reference timing of ignition by the calculated amount to obtain an optimal timing of ignition. 
   Still preferably the storage stores the amount in a map. 
   In accordance with the present invention an optimal timing of ignition obtained by advancing or retarding a reference timing of ignition by a variation introduced to time ignition can be determined from an amount of variation introduced to time ignition that is stored in a map as based on a fuel injection ratio of the in-cylinder and intake manifold injectors. 
   Still preferably, the storage stores the amount divided into a first map applied when the first fuel injection mechanism alone injects the fuel, a second map applied when the second fuel injection mechanism alone injects the fuel, and a third map applied when the first and second fuel injection mechanisms inject the fuel. 
   In accordance with the present invention the fuel injection ratio and the internal combustion engine&#39;s temperature contribute to how an air fuel mixture is formed (e.g., how it is atomized, how homogeneous it is, and the like). When an in-cylinder injector corresponding to one example of the first fuel injection mechanism and an intake manifold injector corresponding to one example of the second fuel injection mechanism bear shares, respectively, of injecting fuel, an amount of variation from a reference timing of ignition is stored in a map divided into a first map applied when the in-cylinder injector alone injects the fuel, a second map applied when the intake manifold injector alone injects the fuel, and a third map applied when the in-cylinder and intake manifold injectors inject the fuel. A map can be selected as based on a fuel injection ratio between the in-cylinder and intake manifold injectors to determine an amount of variation from the reference timing of ignition. Each map can store the amount of variation introduced to time ignition with the internal combustion engine&#39;s temperature serving as a parameter. 
   Still more preferably the first map provides the amount set to provide spark retard. 
   In accordance with the present invention in the first map applied when the first fuel injection mechanism (e.g., an in-cylinder injector) alone injects fuel the fuel injected through the in-cylinder injector (at compression stroke in particular) is atomized satisfactorily despite a cold state and thus mixed with air to form a satisfactorily homogeneous air fuel mixture allowing a high combustion rate. Accordingly, basically, the map stores an amount of variation introduced to time ignition that sets a timing of ignition to be retarded. In particular, such tendency is increased as the internal combustion engine has higher temperature. Accordingly the internal combustion engine&#39;s temperature can be set as a parameter. It should be noted however that while the fuel injected through the in-cylinder injector is injected into the hot cylinder and thus satisfactorily atomized, it is mixed with air before ignition for a short period of time. As such, while the fuel is satisfactorily atomized, the fuel mixed with air may not provide satisfactory homogeneity with the air. Accordingly an amount of variation introduced to time ignition is determined from a relationship between an air fuel mixture&#39;s insufficient homogeneity resulting from a short period of time for mixture that is attributed to the in-cylinder injector and the fuel&#39;s sufficient atomization attributed to the cylinder&#39;s high internal temperature. 
   Still preferably the first map provides the amount set to provide spark advance. 
   In accordance with the present invention in the second map applied when the second fuel injection mechanism (e.g., an intake manifold injector) alone injects fuel the fuel injected through the intake manifold injector is atomized unsatisfactorily in a cold state and thus mixed with air to form an inhomogeneous air fuel mixture contributing to a low combustion rate. Accordingly, basically, the map stores an amount of variation introduced to time ignition that sets a timing of ignition to be advanced. It should be noted however that despite the cold state the fuel injected through the intake manifold injector is mixed with air before ignition for a long period of time. As such, despite its insufficient atomization, the fuel mixed with air may provide satisfactory homogeneity with the air. Accordingly an amount of variation introduced to time ignition is determined from a relationship between an air fuel mixture&#39;s sufficient homogeneity resulting from a long period of time for mixture that is attributed to the intake manifold injector and the fuel&#39;s insufficient atomization attributed to the intake manifold&#39;s low internal temperature. 
   Still preferably the third map provides the amount set to provide spark retard when the first fuel injection mechanism is increased in the ratio. 
   When the first fuel injection mechanism (e.g., an in-cylinder injector) injecting fuel into a hot cylinder despite a cold state and thus allowing the fuel to be satisfactorily atomized and thus mixed with air to form a mixture allowing a high combustion rate has a higher fuel injection ratio, a more sufficient period of time for combustion can be ensured despite an amount of variation set to provide spark retard than when the second fuel injection mechanism (e.g., an intake manifold injector) injecting the fuel into a cold intake manifold in the cold state and thus preventing the fuel from being satisfactorily atomized and mixed with air to form a mixture allowing a high combustion rate has a higher fuel injection ratio. Thus the internal combustion engine having two fuel injection mechanisms that bear shares, respectively, of injecting fuel and provide air fuel mixtures having different conditions, respectively, can achieve an accurately set timing of ignition that can prevent detrimental effects attributed to excessive spark retard and advance. 
   Still preferably the third map provides the amount set to provide spark advance when the second fuel injection mechanism is increased in the ratio. 
   When the second fuel injection mechanism (e.g., an intake manifold injector) injecting fuel into a cold intake manifold in a cold state and thus preventing the fuel from being satisfactorily atomized and mixed with air to form a mixture contributing to a high combustion rate has a higher fuel injection ratio, an amount of variation introduced to time ignition can be provided for spark advance to ensure a more sufficient period of time for combustion (from a timing of ignition to that of starting an exhaust stroke) than when the first fuel injection mechanism (e.g., an in-cylinder injector) injecting the fuel into a hot cylinder despite the cold state and thus allowing the fuel to be satisfactorily atomized and thus mixed with air to form a mixture allowing a high combustion rate has a higher fuel injection ratio. Thus the internal combustion engine having two fuel injection mechanisms that bear shares, respectively, of injecting fuel and provide air fuel mixtures having different conditions, respectively, can achieve an accurately set timing of ignition that can prevent detrimental effects attributed to excessive spark retard and advance. 
   Still preferably the first fuel injection mechanism is an in-cylinder injector and the second fuel injection mechanism is an intake manifold injector. 
   In accordance with the present invention a control apparatus can be provided that calculate an accurate amount of spark advance for an internal combustion engine having separately provided first and second fuel injection mechanisms implemented by an in-cylinder injector and an intake manifold injector to share injecting fuel when they share injecting the fuel in a cold state and a transitional period from the cold state to a warm state. 
   The foregoing and other objects, features, aspects and advantages of the present invention will become more apparent from the following detailed description of the present invention when taken in conjunction with the accompanying drawings. 

   
     BRIEF DESCRIPTION OF THE DRAWINGS 
       FIG. 1  a schematic configuration diagram of an engine system controlled by a control apparatus according to an embodiment of the present invention. 
       FIG. 2  is a flowchart of a program executed by an engine ECU. 
       FIG. 3  shows an example of a map of shared injection. 
       FIG. 4  illustrates how the engine&#39;s operation state varies. 
       FIGS. 5 and 7  show a DI ratio map for a warm state of an engine to which the present control apparatus is suitably applied. 
       FIGS. 6 and 8  show a DI ratio map for a cold state of an engine to which the present control apparatus is suitably applied. 
   

   DESCRIPTION OF THE PREFERRED EMBODIMENTS 
   Hereinafter reference will be made to the drawings to describe the present invention in an embodiment. In the following description identical components are identically denoted. They are also identical in name and function. Note that while the following description is provided exclusively in conjunction with spark advance in a cold state, the present invention is not limited to such advance. The present invention also includes once introducing a spark advance and then a spark retard and introducing a spark retard from a reference timing of ignition. Furthermore, a relationship between a smaller spark advance for a higher ratio of fuel injected through an in-cylinder injector and a significantly large spark advance for a higher ratio of fuel injected through an intake manifold injector, can be inverted. For example, if the performance of an in-cylinder injector  110  as a discrete injector and that of an intake manifold injector  120  as a discrete injector contribute to less sufficient atomization of the fuel injected through in-cylinder  110  than that of the fuel injected through intake manifold injector  120  for the same engine coolant temperature THW, the fuel injection ratio spark advance relationship described above can be inverted. 
     FIG. 1  is a schematic configuration diagram of an engine system that is controlled by an engine ECU (Electronic Control Unit) implementing the control apparatus for an internal combustion engine according to an embodiment of the present invention. In  FIG. 1 , an in-line 4-cylinder gasoline engine is shown, although the application of the present invention is not restricted to such an engine. 
   As shown in  FIG. 1 , the engine  10  includes four cylinders  112 , each connected via a corresponding intake manifold  20  to a common surge tank  30 . Surge tank  30  is connected via an intake duct  40  to an air cleaner  50 . An airflow meter  42  is arranged in intake duct  40 , and a throttle valve  70  driven by an electric motor  60  is also arranged in intake duct  40 . Throttle valve  70  has its degree of opening controlled based on an output signal of an engine ECU  300 , independently from an accelerator pedal  100 . Each cylinder  112  is connected to a common exhaust manifold  80 , which is connected to a three-way catalytic converter  90 . 
   Each cylinder  112  is provided with an in-cylinder injector  110  for injecting fuel into the cylinder and an intake manifold injector  120  for injecting fuel into an intake port or/and an intake manifold. Injectors  110  and  120  are controlled based on output signals from engine ECU  300 . Further, in-cylinder injector  110  of each cylinder is connected to a common fuel delivery pipe  130 . Fuel delivery pipe  130  is connected to a high-pressure fuel pump  150  of an engine-driven type, via a check valve  140  that allows a flow in the direction toward fuel delivery pipe  130 . In the present embodiment, an internal combustion engine having two injectors separately provided is explained, although the present invention is not restricted to such an internal combustion engine. For example, the internal combustion engine may have one injector that can effect both in-cylinder injection and intake manifold injection. 
   As shown in  FIG. 1 , the discharge side of high-pressure fuel pump  150  is connected via an electromagnetic spill valve  152  to the intake side of high-pressure fuel pump  150 . As the degree of opening of electromagnetic spill valve  152  is smaller, the quantity of the fuel supplied from high-pressure fuel pump  150  into fuel delivery pipe  130  increases. When electromagnetic spill valve  152  is fully open, the fuel supply from high-pressure fuel pump  150  to fuel delivery pipe  130  is stopped. Electromagnetic spill valve  152  is controlled based on an output signal of engine ECU  300 . 
   Each intake manifold injector  120  is connected to a common fuel delivery pipe  160  on a low pressure side. Fuel delivery pipe  160  and high-pressure fuel pump  150  are connected via a common fuel pressure regulator  170  to a low-pressure fuel pump  180  of an electric motor-driven type. Further, low-pressure fuel pump  180  is connected via a fuel filter  190  to a fuel tank  200 . Fuel pressure regulator  170  is configured to return a part of the fuel discharged from low-pressure fuel pump  180  back to fuel tank  200  when the pressure of the fuel discharged from low-pressure fuel pump  180  is higher than a preset fuel pressure. This prevents both the pressure of the fuel supplied to intake manifold injector  120  and the pressure of the fuel supplied to high-pressure fuel pump  150  from becoming higher than the above-described preset fuel pressure. 
   Engine ECU  300  is implemented with a digital computer, and includes a ROM (Read Only Memory)  320 , a RAM (Random Access Memory)  330 , a CPU (Central Processing Unit)  340 , an input port  350 , and an output port  360 , which are connected to each other via a bidirectional bus  310 . 
   Airflow meter  42  generates an output voltage that is proportional to an intake air quantity, and the output voltage is input via an A/D converter  370  to input port  350 . A coolant temperature sensor  380  is attached to engine  10 , and generates an output voltage proportional to a coolant temperature of the engine, which is input via an AID converter  390  to input port  350 . 
   A fuel pressure sensor  400  is attached to fuel delivery pipe  130 , and generates an output voltage proportional to a fuel pressure within fuel delivery pipe  130 , which is input via an A/D converter  410  to input port  350 . An air-fuel ratio sensor  420  is attached to an exhaust manifold  80  located upstream of three-way catalytic converter  90 . Air-fuel ratio sensor  420  generates an output voltage proportional to an oxygen concentration within the exhaust gas, which is input via an A/D converter  430  to input port  350 . 
   Air-fuel ratio sensor  420  of the engine system of the present embodiment is a full-range air-fuel ratio sensor (linear air-fuel ratio sensor) that generates an output voltage proportional to the air-fuel ratio of the air-fuel mixture burned in engine  10 . As air-fuel ratio sensor  420 , an O 2  sensor may be employed, which detects, in an on/off manner, whether the air-fuel ratio of the air-fuel mixture burned in engine  10  is rich or lean with respect to a theoretical air-fuel ratio. 
   Accelerator pedal  100  is connected with an accelerator pedal position sensor  440  that generates an output voltage proportional to the degree of press down of accelerator pedal  100 , which is input via an A/D converter  450  to input port  350 . Further, an engine speed sensor  460  generating an output pulse representing the engine speed is connected to input port  350 . ROM  320  of engine ECU  300  prestores, in the form of a map, values of fuel injection quantity that are set in association with operation states based on the engine load factor and the engine speed obtained by the above-described accelerator pedal position sensor  440  and engine speed sensor  460 , and correction values thereof set based on the engine coolant temperature. 
   With reference to the flowchart of  FIG. 2  engine ECU  300  of  FIG. 1  executes a program having a structure for control, as described hereinafter. 
   In step (S)  100  engine ECU  300  employs a map as shown in  FIG. 3  to calculate an injection ratio of in-cylinder injector  110 . Hereinafter this ratio will be referred to as “DI ratio r,” wherein 0≦r≦1. The map used to calculate the ratio will be described later. 
   In S 100  engine ECU  300  determines whether DI ratio r is 1, 0, or larger than 0 and smaller than 1. If DI ratio r is 1 (r=1.0 in S 110 ) the process proceeds to S 120 . If DI ratio r is 0 (r=0 in S 110 ) the process proceeds to S 130 . If DI ratio r is larger than 0 and smaller than 1 (0&lt;r&lt;1 in S 110 ) the process proceeds to S 140 . 
   In S 120  engine ECU  300  calculates an amount of cold state spark advance corresponding to that of spark advance for correction in a cold state when in-cylinder injector  110  alone injects fuel. This is done for example by employing a function f(1) to calculate an amount of cold state spark advance=f(1)(THW). Note that “THW” represents the temperature of a coolant of engine  10  as detected by coolant temperature sensor  380 . 
   Note that when the fuel injected through in-cylinder injector  110  and that injected through intake manifold injector  120  are mixed with air they provide air fuel mixtures having significantly different conditions. In particular, in a cold state when the fuel injected through in-cylinder injector  110  into a hot cylinder and that injected through intake manifold injector  120  into a cold intake manifold are mixed with air they provide air fuel mixtures having significantly different conditions. The fuel injected through intake manifold injector  120  is atomized insufficiently and hence mixed with air to provide an air fuel mixture having an unsatisfactory (or inhomogeneous) condition resulting in a low combustion rate. In contrast, the fuel injected through in-cylinder injector  110  is satisfactorily atomized and hence mixed with air to provide an air fuel mixture having a satisfactory (or homogeneous) condition allowing a high combustion rate. 
   Function f(1)(THW) is a function applied for DI ratio r=0 (i.e., when in-cylinder injector  110  alone injects fuel). It is set to allow an amount of cold state spark advance to be minimized (including providing a spark retard from a reference timing of ignition.), since the fuel injected through in-cylinder injector  110  and air can form a satisfactorily homogenous air fuel mixture allowing an increased combustion rate, and if ignition is timed to be maximally retarded a sufficient period of time for combustion can still be ensured. 
   Despite that, the period after in-cylinder injector  110  injects the fuel before ignition is short. As such, while the fuel is satisfactorily atomized, there is a tendency that a sufficient period of time is not provided to ensure homogeneity. Thus a short period of time elapsing after in-cylinder injector  110  injects the fuel and before ignition affects an air fuel mixture to be unsatisfactorily homogeneous. Accordingly, setting function f(1)(THW) while considering such a period of time elapsing after fuel is injected and before ignition, can also be considered. 
   In S 130  engine ECU  300  calculates an amount of cold state spark advance corresponding to that of spark advance for correction in the cold state when intake manifold injector  120  alone injects fuel. This is done for example by employing a function f(2) to calculate an amount of cold state spark advance=f(2)(THW). 
   Function f(2)(THW) is a function applied for DI ratio r=0 (i.e., when intake manifold injector  120  alone injects fuel). It is set to allow an amount of cold state spark advance to be maximized, since the fuel injected through intake manifold injector  120  and air form an inhomogeneous air fuel mixture contributing to a decreased combustion rate, and accordingly ignition is timed to be maximally advanced to ensure a sufficient period of time for combustion. 
   Despite that, the period after intake manifold injector  120  injects the fuel and before ignition is long. As such, while the fuel is unsatisfactorily atomized, there is a tendency that a sufficient period of time is provided to ensure homogeneity. As such, a long period of time elapsing after intake manifold injector  120  injects the fuel and before ignition affects an air fuel mixture to be satisfactorily homogeneous. Accordingly, setting function f(2)(THW) while considering such a period of time elapsing after fuel is injected and before ignition, can also be considered. 
   In S 140  engine ECU  300  calculates an amount of cold state spark advance corresponding to that of spark advance for correction in a cold state when in-cylinder and intake manifold injectors  110  and  120  bear shares, respectively, of injecting fuel. This is done for example by employing a function f(3) to calculate an amount of cold state spark advance=f(3)(THW, r). Note that “r” represents a DI ratio. 
   In S 150  engine ECU  300  calculates a timing of ignition for example by employing a function g to calculate a timing of ignition=g (an amount of cold state spark advance). 
   Reference will now be made to  FIG. 3  to describe an injection ratio (0≦DI ratio r≦1) of in-cylinder injector  110  with an engine speed NE and a load factor KL of engine  10  serving as parameters. 
   In a low engine speed and high load range the fuel injected through in-cylinder injector  10  is insufficiently mixed with air, and in the combustion chamber the air fuel mixture tends to be inhomogeneous and thus provide unstable combustion. Accordingly, for this range, DI ratio r is reduced to increase an injection ratio (1−r) of intake manifold injector  120  to sufficiently mix the air fuel mixture before it is introduced into the combustion chamber. 
   In a high engine speed and low load range the air fuel mixture injected through in-cylinder injector  10  is readily homogenized. Accordingly, DI ratio r is increased. The fuel injected through in-cylinder injector  10  is vaporized within the combustion chamber involving latent heat of vaporization (by absorbing heat from the combustion chamber). Accordingly at the compression side the air fuel mixture is decreased in temperature and improved antiknock performance is provided. Furthermore, as the combustion chamber is decreased in temperature, improved intake efficiency can be achieved and high power output expected. Furthermore, in-cylinder injector  110  can have its end, exposed in the combustion chamber, cooled by the fuel and thus have its injection hole prevented from having deposit adhering thereto. 
   As based on the configuration and flowchart as described above, engine  10  in the present embodiment operates as described hereinafter. Note that in the following description “if the engine&#39;s coolant varies in temperature” and other similar expressions indicate a transitional period from a cold state to a warm state. 
   No Variation in DI Ratio R and Variation Present in Temperature of Coolant for Engine 
   When engine  10  starts, normally the coolant increases in temperature. More specifically, in  FIG. 4 , the coolant increases in temperature from a temperature TH( 1 ) corresponding to a point A to a temperature TH( 2 ) corresponding to a point B. The DI ratio is calculated (S 100 ) and if DI ratio r is not found to have varied (e.g., r=0.7) a decision is made that it is larger than 0 and smaller than 1 (0&lt;r&lt;1 in S 110 ) and function f(3) is accordingly used to calculate an amount of cold state spark advance by f(3) (THW, r) (S 140 ). 
   In  FIG. 4 , for point A, by f(3) (TH( 1 ), r), wherein r=0.7, an amount of cold state spark advance is calculated as a spark advance for correction ( 1 ). With the amount of cold state spark advance set at the spark advance for correction ( 1 ), engine  10  is operated, and temperature THW increases from TH( 1 ) to TH( 2 ) to reach point B. For point B, by f(3) (TH( 2 ), r), wherein r=0.7, an amount of cold state spark advance is calculated as a spark advance for correction ( 2 ). In other words, an amount of spark advance for correction is reduced from the spark advance for correction ( 1 ) to the spark advance for correction ( 2 ) by a variation in amount of spark advance for correction, which is provided by the spark advance for correction ( 1 ) minus the spark advance for correction ( 2 ). 
   Variation Present in DI Ratio R and No Variation in Temperature of Coolant for Engine 
   While engine  10  is started, the coolant may not vary depending on the vehicle&#39;s surrounding (temperature in particular). If in such a case the engine  10  operation state varies and DI ratio r drops from 0.7, i.e., in  FIG. 4 , while temperature TH( 1 ) corresponding to point A is held, a point C allowing DI ratio r smaller than 0.7 is attained (or it may be vice versa). The DI ratio is calculated and if DI ratio r is found to have varied (for example from 0.7 to 0.5) a decision is made that DI ratio r is still larger than 0 and smaller than 1 (0&lt;r&lt;1 in S 100 ), and function f(3) is employed to calculate an amount of cold state spark advance by f(3) (THW, r) (S 140 ). 
   In  FIG. 4 , for point A, by f(3) (TH( 1 ), r), wherein r=0.7, an amount of cold state spark advance is calculated. In this condition engine  10  is operated, and while temperature THW is held at TH( 1 ), DI ratio r decreases to reach point C. For point C, by f(3) (TH( 1 ), r), wherein r=0.5, an amount of cold state spark advance is calculated. More specifically, a spark advance is introduced by a variation in amount of spark advance for correction. This indicates that a larger spark advance is introduced as the port&#39;s temperature is lower than the cylinder&#39;s internal temperature and the fuel injected through intake manifold injector  120  is hard to atomize. 
   Variation Present in DI ratio R and Variation Present in Temperature of Coolant for Engine 
   When engine  10  is started the coolant&#39;s temperature and DI ratio r may both vary. In such a case, in  FIG. 4  point A corresponding to temperature TH( 1 ) and DI ratio r=0.7 transitions to a point D corresponding to temperature TH( 2 ) higher than TH( 1 ) and a DI ratio r smaller than 0.7. The DI ratio is calculated and if DI ratio r is found to have varied (for example from 0.7 to 0.5) a decision is still made that DI ratio r is larger than 0 and smaller than 1 (0&lt;r&lt;1 in S 100 ), and function f(3) is employed to calculate an amount of cold state spark advance by f(3) (THW, r) (S 140 ). 
   In  FIG. 4 , for point A, by f(3) (TH( 1 ), r), wherein r=0.7, an amount of cold state spark advance is calculated. In this condition engine  10  is operated, and while temperature THW changes from TH( 1 ) to TH( 2 ) the DI ratio also decreases to reach point D. For point D, by f(3) (TH( 2 ), r), wherein r=0.5, an amount of cold state spark advance is calculated. More specifically, a timing of ignition is varied by a variation in amount of spark advance for correction. This indicates that when a DI ratio is neither 0 nor 1 an amount of cold state spark advance is calculated by a function of the coolant&#39;s temperature and DI ratio r, and a variation in amount of spark advance for correction also depends on those of the coolant in temperature and DI ratio r, respectively. 
   Thus in a cold state and a transitional period from the cold state to a warm state when an in-cylinder injector and an intake manifold injector bear shares, respectively, of injecting fuel, not only temperature THW of the coolant of the engine but DI ratio r is also used to calculate an amount of cold state spark advance. If the cylinder&#39;s interior and the port are different in temperature and thus have fuel therein atomized differently an accurate spark advance can be provided to combust the fuel satisfactorily. 
   Furthermore, as will be described hereinafter, three maps may be stored in engine ECU  300  at ROM  320 , RAM  340  or the like. 
   A first map is set as a map applied for DI ratio r=1 (i.e., when in-cylinder injector  110  alone injects fuel) for a reference timing of ignition, an amount of variation introduced to time ignition, and the like that time ignition to be maximally retarded. The cylinder receiving fuel injected from in-cylinder injector  110  has high internal temperature despite a cold state. As such, the injected fuel can be atomized satisfactorily and thus mixed with air to provide a satisfactorily homogeneous air fuel mixture allowing an increased combustion rate. As such, if spark retard is introduced a sufficient period of time for combustion can still be ensured. Furthermore, as combustion rate is not a factor significantly limiting a timing of ignition, other factors (e.g., catalyst warm up, exhaust gas purification, and the like) can be considered in further retarding or advancing the reference timing of ignition. 
   A second map is set as a map applied for DI ratio r=0 (i.e., when intake manifold injector  120  alone injects fuel) for a reference timing of ignition, an amount of variation introduced to time ignition, and the like that time ignition to be maximally advanced. In a cold state the intake manifold receiving fuel injected from intake manifold injector  120  has low temperature. As such, the injected fuel is atomized unsatisfactorily and thus mixed with air to provide an inhomogeneous air fuel mixture contributing to a decreased combustion rate. Accordingly, a faster timing of ignition must be provided to ensure a sufficient period of time for combustion. 
   A third map is set as a map applied for a DI ratio of 0&lt;r&lt;1 (i.e., when in-cylinder injector  110  and intake manifold injector  120  bear shares, respectively, of injecting fuel) for a reference timing of ignition, an amount of variation introduced to time ignition, and the like that time ignition to be more retarded for higher DI ratio r. As DI ratio r increases, the fuel is atomized more satisfactorily and thus mixed with air to provide a more satisfactorily homogeneous air fuel mixture allowing a higher combustion rate. Accordingly, spark retard is introduced. 
   Engine ECU  300  prepares three maps for such reference timings of ignition, amounts of variation introduced to time ignition, and the like, and in accordance with a ratio of in-cylinder injector  110  bearing a share of injecting fuel, or DI ratio r, selects one of the maps to switch a map of a reference timing of ignition, a map of an amount of variation introduced to time ignition, and the like. In accordance with the selected map engine ECU  300  calculates a reference timing of ignition. In particular, the third map provides a reference timing of ignition, an amount of variation introduced to time ignition, and the like varied by DI ratio r. Accordingly, not only the map but a function interpolating an intermediate portion set in the map may also be previously calculated and stored, and used to provide interpolation. 
   Thus one of the three maps (DI ratio: r=1, r=0, and 0&lt;r&lt;1) can be selected by DI ratio r and used to calculate a reference timing of ignition, an amount of variation introduced to time ignition, and the like. An appropriate reference timing of ignition corresponding to DI ratio r can thus be calculated and set, and detrimental effects of excessive spark retard and advance can be prevented. 
   Engine ( 1 ) Suitable for Application of the Control Apparatus 
   An engine ( 1 ) suitable for application of the control apparatus in the present embodiment will be described hereinafter. 
   Referring to  FIGS. 5 and 6 , maps each indicating a fuel injection ratio between in-cylinder injector  10  and intake manifold injector  120 , identified as information associated with an operation state of engine  10 , will now be described. Herein, the fuel injection ratio between the two injectors will also be expressed as a ratio of the quantity of the fuel injected from in-cylinder injector  10  to the total quantity of the fuel injected, which is referred to as the “fuel injection ratio of in-cylinder injector  110 ”, or, a “DI (Direct Injection) ratio (r)”. The maps are stored in ROM  320  of engine ECU  300 .  FIG. 5  shows the map for the warm state of engine  10 , and  FIG. 6  shows the map for the cold state of engine  10 . 
   In the maps shown in  FIGS. 5 and 6 , with the horizontal axis representing an engine speed of engine  10  and the vertical axis representing a load factor, the fuel injection ratio of in-cylinder injector  110 , or the DI ratio r, is expressed in percentage. 
   As shown in  FIGS. 5 and 6 , the DI ratio r is set for each operation region that is determined by the engine speed and the load factor of engine  10 . “DI RATIO r=100%” represents the region where fuel injection is carried out using only in-cylinder injector  110 , and “DI RATIO r=0%” represents the region where fuel injection is carried out using only intake manifold injector  120 . “DI RATIO r≠0%”, “DI RATIO r≠100%” and “0%&lt;DI RATIO r&lt;100%” each represent the region where fuel injection is carried out using both in-cylinder injector  110  and intake manifold injector  120 . Generally, in-cylinder injector  110  contributes to an increase of output performance, while intake manifold injector  120  contributes to uniformity of the air-fuel mixture. These two kinds of injectors having different characteristics are appropriately selected depending on the engine speed and the load factor of engine  10 , so that only homogeneous combustion is conducted in the normal operation state of engine  10  (other than the abnormal operation state such as a catalyst warm-up state during idling, for example). 
   Further, as shown in  FIGS. 5 and 6 , the fuel injection ratio between in-cylinder injector  110  and intake manifold injector  120  is defined as the DI ratio r, individually in the maps for the warm state and the cold state of the engine. The maps are configured to indicate different control regions of in-cylinder injector  110  and intake manifold injector  120  as the temperature of engine  10  changes. When the temperature of engine  10  detected is equal to or higher than a predetermined temperature threshold value, the map for the warm state shown in  FIG. 5  is selected; otherwise, the map for the cold state shown in  FIG. 6  is selected. One or both of in-cylinder injector  110  and intake manifold injector  120  are controlled based on the selected map and according to the engine speed and the load factor of engine  10 . 
   The engine speed and the load factor of engine  10  set in  FIGS. 5 and 6  will now be described. In  FIG. 5 , NE( 1 ) is set to 2500 rpm to 2700 rpm, KL( 1 ) is set to 30% to 50%, and KL( 2 ) is set to 60% to 90%. In  FIG. 6 , NE( 3 ) is set to 2900 rpm to 3100 rpm. That is, NE( 1 )&lt;NE( 3 ). NE( 2 ) in  FIG. 5  as well as KL( 3 ) and KL( 4 ) in  FIG. 6  are also set as appropriate. 
   When comparing  FIG. 5  and  FIG. 6 , NE( 3 ) of the map for the cold state shown in  FIG. 6  is greater than NE( 1 ) of the map for the warm state shown in  FIG. 5 . This shows that, as the temperature of engine  10  is lower, the control region of intake manifold injector  120  is expanded to include the region of higher engine speed. That is, when engine  10  is cold, deposits are unlikely to accumulate in the injection hole of in-cylinder injector  110  (even if the fuel is not injected from in-cylinder injector  110 ). Thus, the region where the fuel injection is to be carried out using intake manifold injector  120  can be expanded, to thereby improve homogeneity. 
   When comparing  FIG. 5  and  FIG. 6 , “DI RATIO r=100%” holds in the region where the engine speed of engine  10  is equal to or higher than NE( 1 ) in the map for the warm state, and in the region where the engine speed is NE( 3 ) or higher in the map for the cold state. Further, “DI RATIO r=100%” holds in the region where the load factor is KL( 2 ) or greater in the map for the warm state, and in the region where the load factor is KL( 4 ) or greater in the map for the cold state. This means that fuel injection is carried out using only in-cylinder injector  110  in the region where the engine speed is at a predetermined high level, and that fuel injection is carried out using only in-cylinder injector  110  in the region where the engine load is at a predetermined high level, since for the high speed region and the low load region the engine  10  speed and load are high and a large quantity of air is intaken, and in-cylinder injector  110  can singly be used to inject fuel to provide a homogeneous air fuel mixture. In this case, the fuel injected from in-cylinder injector  110  is atomized within the combustion chamber involving latent heat of vaporization (by absorbing heat from the combustion chamber). Accordingly, the temperature of the air-fuel mixture is decreased at the compression side, and thus, the antiknock performance improves. Further, with the temperature of the combustion chamber decreased, intake efficiency improves, leading to high power output. 
   In the map for the warm state in  FIG. 5 , fuel injection is also carried out using only in-cylinder injector  110  when the load factor is KL( 1 ) or less. This shows that in-cylinder injector  110  alone is used in a predetermined low load region when the temperature of engine  10  is high. When engine  10  is in the warm state, deposits are likely to accumulate in the injection hole of in-cylinder injector  110 . However, when fuel injection is carried out using in-cylinder injector  110 , the temperature of the injection hole can be lowered, whereby accumulation of deposits is prevented. Further, clogging of in-cylinder injector  110  may be prevented while ensuring a minimum fuel injection quantity thereof. Thus, in-cylinder injector  110  alone is used in the relevant region. 
   When comparing  FIG. 5  and  FIG. 6 , there is a region of “DI RATIO r=0%” only in the map for the cold state in  FIG. 6 . This shows that fuel injection is carried out using only intake manifold injector  120  in a predetermined low load region (KL( 3 ) or less) when the temperature of engine  10  is low. When engine  10  is cold and low in load and the intake air quantity is small, atomization of the fuel is unlikely to occur. In such a region, it is difficult to ensure favorable combustion with the fuel injection from in-cylinder injector  110 . Further, particularly in the low-load and low-speed region, high power output using in-cylinder injector  110  is unnecessary. Accordingly, fuel injection is carried out using intake manifold injector  120  alone, rather than using in-cylinder injector  110 , in the relevant region. 
   Further, in an operation other than the normal operation, i.e., in the catalyst warm-up state at idle of engine  10  (abnormal operation state), in-cylinder injector  110  is controlled to carry out stratified charge combustion. By causing the stratified charge combustion during the catalyst warm-up operation, warming up of the catalyst is promoted, and exhaust emission is thus improved. 
   Engine ( 2 ) Suitable for Application of the Control Apparatus 
   An engine ( 2 ) suitable for application of the control apparatus in the present embodiment will be described hereinafter. In the following description of engine ( 2 ) the same description as that of engine ( 1 ) will not be repeated. 
   Referring to  FIGS. 7 and 8 , maps each indicating a fuel injection ratio between in-cylinder injector  110  and intake manifold injector  120 , identified as information associated with an operation state of engine  10 , will now be described. The maps are stored in ROM  320  of engine ECU  300 .  FIG. 7  shows the map for the warm state of engine  10 , and  FIG. 8  shows the map for the cold state of engine  10 . 
   When comparing  FIG. 7  and  FIG. 8 , the figures differ from  FIGS. 5 and 6 , as follows: “DI RATIO r=100%” holds in the region where the engine speed of engine  10  is equal to or higher than NE( 1 ) in the map for the warm state, and in the region where the engine speed is NE( 3 ) or higher in the map for the cold state. Further, except for the low-speed region, “DI RATIO r=100%” holds in the region where the load factor is KL( 2 ) or greater in the map for the warm state, and in the region where the load factor is KL( 4 ) or greater in the map for the cold state. This means that fuel injection is carried out using only in-cylinder injector  110  in the region where the engine speed is at a predetermined high level, and that fuel injection is often carried out using only in-cylinder injector  110  in the region where the engine load is at a predetermined high level. However, in the low-speed and high-load region, mixing of an air-fuel mixture formed by the fuel injected from in-cylinder injector  110  is poor, and such inhomogeneous air-fuel mixture within the combustion chamber may lead to unstable combustion. Accordingly, the fuel injection ratio of in-cylinder injector  110  is increased as the engine speed increases where such a problem is unlikely to occur, whereas the fuel injection ratio of in-cylinder injector  110  is decreased as the engine load increases where such a problem is likely to occur. These changes in the fuel injection ratio of in-cylinder injector  110 , or, the DI ratio r, are shown by crisscross arrows in  FIGS. 7 and 8 . In this manner, variation in output torque of the engine attributable to the unstable combustion can be suppressed. It is noted that these measures are approximately equivalent to the measures to decrease the fuel injection ratio of in-cylinder injector  110  as the state of the engine moves toward the predetermined low speed region, or to increase the fuel injection ratio of in-cylinder injector  110  as the engine state moves toward the predetermined low load region. Further, except for the relevant region (indicated by the crisscross arrows in  FIGS. 7  and  8 ), in the region where fuel injection is carried out using only in-cylinder injector  110  (on the high speed side and on the low load side), a homogeneous air-fuel mixture is readily obtained even when the fuel injection is carried out using only in-cylinder injector  110 . In this case, the fuel injected from in-cylinder injector  110  is atomized within the combustion chamber involving latent heat of vaporization (by absorbing heat from the combustion chamber). Accordingly, the temperature of the air-fuel mixture is decreased at the compression side, and thus, the antiknock performance improves. Further, with the temperature of the combustion chamber decreased, intake efficiency improves, leading to high power output. 
   In engine  10  described with reference to  FIGS. 5–8 , homogeneous combustion is achieved by setting the fuel injection timing of in-cylinder injector  110  in the intake stroke, while stratified charge combustion is achieved by setting it in the compression stroke. That is, when the fuel injection timing of in-cylinder injector  110  is set in the compression stroke, a rich air-fuel mixture can be located locally around the spark plug, so that a lean air-fuel mixture in the combustion chamber as a whole is ignited to realize the stratified charge combustion. Even if the fuel injection timing of in-cylinder injector  110  is set in the intake stroke, stratified charge combustion can be realized if it is possible to locate a rich air-fuel mixture locally around the spark plug. 
   As used herein, the stratified charge combustion includes both the stratified charge combustion and semi-stratified charge combustion. In the semi-stratified charge combustion, intake manifold injector  120  injects fuel in the intake stroke to generate a lean and homogeneous air-fuel mixture in the whole combustion chamber, and then in-cylinder injector  110  injects fuel in the compression stroke to generate a rich air-fuel mixture around the spark plug, so as to improve the combustion state. Such semi-stratified charge combustion is preferable in the catalyst warm-up operation for the following reasons. In the catalyst warm-up operation, it is necessary to considerably retard the ignition timing and maintain favorable combustion state (idling state) so as to cause a high-temperature combustion gas to reach the catalyst. Further, a certain quantity of fuel needs to be supplied. If the stratified charge combustion is employed to satisfy these requirements, the quantity of the fuel will be insufficient. With the homogeneous combustion, the retarded amount for the purpose of maintaining favorable combustion is small compared to the case of stratified charge combustion. For these reasons, the above-described semi-stratified charge combustion is preferably employed in the catalyst warm-up operation, although either of stratified charge combustion and semi-stratified charge combustion may be employed. 
   Furthermore in the engine described with reference to  FIGS. 5–8  preferably in-cylinder injector  110  is timed to inject fuel at the compression stroke for the following reason, although in engine  10  described above, the fuel injection timing of in-cylinder injector  110  is set in the intake stroke in a basic region corresponding to the almost entire region (herein, the basic region refers to the region other than the region where semi-stratified charge combustion is conducted by causing intake manifold injector  120  to inject the fuel in the intake stroke and causing in-cylinder injector  110  to inject the fuel in the compression stroke, which is conducted only in the catalyst warm-up state). The fuel injection timing of in-cylinder injector  110 , however, may be set temporarily in the compression stroke for the purpose of stabilizing combustion, for the following reasons. 
   When the fuel injection timing of in-cylinder injector  110  is set in the compression stroke, the air-fuel mixture is cooled by the injected fuel while the temperature in the cylinder is relatively high. This improves the cooling effect and, hence, the antiknock performance. Further, when the fuel injection timing of in-cylinder injector  110  is set in the compression stroke, the time from the fuel injection to the ignition is short, which ensures strong penetration of the injected fuel, so that the combustion rate increases. The improvement in antiknock performance and the increase in combustion rate can prevent variation in combustion, and thus, combustion stability is improved. 
   Note that in the above described flowchart at S 150  whenever the flowchart is executed a reference timing of ignition may be calculated from the engine  10  operation state and function g correcting the reference timing of ignition by an amount of cold state spark advance may be used to calculate a timing of ignition. 
   Furthermore, irrespectively of the engine  10  temperature (i.e., in either a warm state or a cold state) when idling is off (i.e., an idle switch is off, the accelerator pedal is pressed) the  FIG. 5  or  7  map for a warm state may be used. (Regardless of cold or warm state, in-cylinder injector  110  is used for a low load range.) 
   Although the present invention has been described and illustrated in detail, it is clearly understood that the same is by way of illustration and example only and is not to be taken by way of limitation, the spirit and scope of the present invention being limited only by the terms of the appended claims.