JP4793301B2 - Spark ignition gasoline engine - Google Patents

Description

  The present invention relates to a spark ignition gasoline engine.

In general, as disclosed in Non-Patent Document 1, a spark-ignition gasoline engine theoretically follows an Otto Cycle, and its theoretical thermal efficiency is η _th.
η _th = 1− (1 / ε ^κ−1 ) (1)
(Where ε is a compression ratio and κ is a specific heat ratio).

  As is clear from the equation (1), the theoretical thermal efficiency of the spark-ignition gasoline engine (therefore, shown in the figure, net thermal efficiency) is improved to a certain level when the compression ratio is high. In this regard, Non-Patent Document 1 describes the theoretical thermal efficiency of various compression ratios (8 ≦ ε ≦ 20) when a spark ignition gasoline engine is operated at 2000 rpm with the throttle fully open (so-called WOT: Wide-Open Throttle). Research that examines changes is introduced. According to the description, the theoretical thermal efficiency and mean effective pressure (MEP) increase proportionally until the compression ratio reaches around 17, and then remain flat.

  Against the background of the above research results, attempts have been made to commercialize high compression ratio engines.

  However, in a spark-ignition engine with a high compression ratio, output reduction due to knocking in a high-load operation region including the throttle full open region is inevitable.

  In this regard, as a general countermeasure against knocking, ignition retard that retards the ignition timing is widely known. However, in the high load operation region including the throttle fully open region, it has been considered that knocking avoidance by ignition retard greatly reduces the output and greatly impairs the merchantability.

  FIG. 1 is a graph showing an example of ignition retard during high load operation.

  For example, as shown in FIG. 1, with the compression ratio (ε = 11) widely used in ordinary engines, knocking does not occur when the ignition timing is set to 4 ° before compression top dead center, but the turbocharger is When used, knocking occurs even when the ignition timing is 4 ° before compression top dead center when the compression ratio is high (ε = 10). Therefore, it has been considered that a large ignition timing retard is necessary to employ a high compression ratio. This means that when the compression ratio is increased to about 10, the output decrease due to the ignition timing retard for preventing knocking surpasses the output increase due to the improvement of the compression ratio, and the output is greatly decreased. For this reason, conventionally, in consideration of a decrease in output due to ignition timing retard, when using a turbocharger in a high load operation region including the throttle full open region, the compression ratio = 10 is set as the limit of the high compression ratio, No higher compression ratio was used.

  In view of this, in the high load operation region including the throttle fully open region, a method of reducing the effective compression ratio by using a so-called Atkinson cycle or Miller cycle is known. However, if the closing timing of the intake valve is changed during high load operation and the effective compression ratio is lowered, the fresh air is lost and the pressure is reduced and the charging efficiency is lowered and the output is reduced in the intake stroke.

Therefore, a technique for reducing the geometric compression ratio of the engine in a high load operation region including the throttle full open region is also known. For example, Patent Documents 1 and 2 disclose a technique in which a variable compression ratio mechanism for changing a geometric compression ratio is provided in an engine, and the geometric compression ratio is changed according to an operation state.
Japanese Patent Laying-Open No. 2005-076579 JP 2005-146991 A “Internal Combustion Engine Fundamentals” by John B. Heywood

  In the above-described conventional techniques, in all throttle areas, the compression ratio is reduced to avoid knocking. For this reason, the approach to high compression ratios in spark-ignited gasoline engines has been forced to choose between sacrificing output and cost.

  Moreover, providing a mechanism for changing the geometric compression ratio as disclosed in Patent Documents 1 and 2 complicates the engine and increases the cost.

  The present invention has been made in view of the above problems, and provides a spark ignition gasoline engine having both low cost and high output even in a high load operation region (especially a throttle full open region) in a low speed region. It is an issue.

  As a result of diligent research, the present inventor has found that after the compression top dead center in an engine having a high compression ratio (ε = 12 or more) such that the ignition timing determined from the knocking limit when using a supercharger is after the compression top dead center. In this case, the cool flame reaction in the cylinder becomes prominent, and by this cool flame reaction, the output increase due to the compression ratio improvement far exceeds the output decrease due to the ignition timing retard to prevent knocking. The present invention has been completed.

That is, in order to solve the above-mentioned problems, the present invention relates to an engine body having a geometric compression ratio set to 13 or more and a cylinder of the engine body in a four-cycle spark ignition gasoline engine having at least a spark plug. An intake valve and an exhaust valve that are respectively provided to the connected intake port and exhaust port, and open and close the corresponding ports; a supercharger that is provided in an intake passage for introducing fresh air into the intake port; Based on the detection of the operating state detecting means, the operating state detecting means for detecting the operating state, the adjustment control of the effective compression ratio by at least the ignition timing adjustment control of the ignition plug and the intake valve closing timing adjustment control, and the excess Control means for executing drive control of the feeder, and the control means has a small operating range of the engine body. In the case of a high load operation region including a throttle full open region in the low speed region, the intake valve closing timing is adjusted so as to maintain the effective compression ratio obtained at the intake valve closing timing defined by the valve lift of 1 mm to 12 or more. In addition, the ignition timing is retarded within a predetermined period after the compression top dead center to cause a cold flame reaction after the compression top dead center has elapsed, and then the air-fuel mixture is burned by flame propagation based on spark ignition. This is a spark ignition type gasoline engine. In this mode, normally, in order to prevent knocking, an effective compression ratio obtained at the intake valve closing timing defined by the valve lift of 1 mm in an operation region where it is considered that a significant ignition timing retard is necessary. The engine body is operated while maintaining a high torque and fuel consumption of 12 or more. That is, when the effective compression ratio is 12 or more and the ignition timing retarded to avoid knocking is set after the compression top dead center, the cool flame in the cylinder after the compression top dead center has elapsed. The reaction becomes remarkable, and the combustion process after the elapse of the compression top dead center becomes multistage ignition. As a result, the heat generation rate (dQ / dθ) can be maintained while reducing time loss, and sufficient torque can be obtained. It becomes possible. In addition, by maintaining such a heat generation rate, the amount of retard can be reduced as much as possible. On the other hand, in the region where the cold flame reaction occurs, the number of moles increases, and as a result, the in-cylinder temperature does not increase as much as the pressure increases. In addition, the cold flame reaction occurs on the center side of the combustion chamber, and since there is little generation of end gas, an increase in the in-cylinder temperature is also suppressed. Under such temperature conditions, formaldehyde (HCHO) is generated, and this formaldehyde promotes consumption of OH radicals that cause knocking, and autoignition is also suppressed from this point. By constructing such a knocking suppression mechanism in a high compression ratio in the high load operation region including the throttle full open region at least in the low speed region, the reduction in output due to the retard of the ignition timing is compensated by the thermal efficiency improvement due to the cold flame reaction. It is also possible to obtain as much fuel efficiency as a diesel engine without sacrificing output. Further, since the effective compression ratio is determined by the closing timing adjustment control of the intake valve, it is not necessary to use a complicated mechanism for changing the geometric compression ratio.

  In the spark ignition gasoline engine described above, the engine body is preferably operated using fuel having an octane number of 96 RON or more. In that case, in the high load operation region including at least the throttle fully open region in the low speed region, the effective compression ratio is set to 12 or more, and the ignition timing is retarded within a predetermined period, so that the cooling in the cylinder is most effectively performed. High torque can be obtained using the flame reaction. As will be described in detail later, when fuel of 96 RON or more is injected, the compression ratio is 13 or more and the activation energy causing the cold flame reaction is increased, and the amount of heat generated by the cold flame reaction is improved by the ignition retard, The torque can be increased. In an engine using a supercharger, by reducing the compression ratio by 1, it is possible to increase the amount of heat generated by the cold flame reaction and increase the torque by ignition retard without causing self-ignition. It is.

  In the spark ignition gasoline engine described above, the upper limit of the geometric compression ratio of the engine body is preferably 15. In such a case, even if a high effective compression ratio is maintained in a situation where self-ignition is likely to occur, such as when the intake air temperature is low and the engine is warm, or when the engine is warm, the pre-ignition etc. Occurrence can be prevented.

  In the spark ignition gasoline engine described above, the engine body is operated using a fuel having an octane number of 100 RON or more, and the upper limit of the geometric compression ratio of the engine body is 15.5. In such a case, prevent pre-ignition from occurring even if a high effective compression ratio is maintained under conditions where auto-ignition is likely to occur, such as when the intake air temperature is high or when the engine is warm. Can do.

Another aspect of the present invention is an engine that is operated using a fuel having a geometric compression ratio of 12.5 or higher and an octane number of 91 RON or higher in a four-cycle spark ignition gasoline engine having at least a spark plug. Provided in a main body, an intake port and an exhaust port connected to a cylinder of the engine body, respectively, and provided in an intake valve and an exhaust valve for opening and closing the corresponding ports, and an intake passage for introducing fresh air into the intake port Based on the detection of the supercharger, the operating state of the engine body, and the detection of the operating state detector, at least the ignition timing adjustment control of the ignition plug and the intake valve closing timing adjustment control Control means for executing adjustment control of the effective compression ratio and drive control of the supercharger, and the control means When the operating range of the gin main body is a high load operating range including at least the throttle fully open range in the low speed range, the effective compression ratio obtained at the intake valve closing timing defined by the valve lift of 1 mm is maintained at 11.5 or more. By adjusting the intake valve closing timing and retarding the ignition timing within a predetermined period after compression top dead center, a cold flame reaction occurs after the compression top dead center has elapsed, and then flame propagation based on spark ignition This is a spark ignition type gasoline engine characterized in that the air-fuel mixture is burned by the above. In this aspect, even when a fuel having a relatively low octane number is used, a high flame is obtained by effectively utilizing the in-cylinder cool flame reaction in a high load operation region including at least the throttle fully open region in the low speed region. be able to.

  In a spark ignition gasoline engine operated using a fuel of 91 RON or more, the upper limit of the geometric compression ratio of the engine body is preferably 14.5. In such a case, prevent pre-ignition from occurring even if a high effective compression ratio is maintained under conditions where auto-ignition is likely to occur, such as when the intake air temperature is high or when the engine is warm. Can do.

In the aspect in which the geometric compression ratio is set to 13 or more, the control means reduces the effective compression ratio to less than 12 and causes compression top dead when the operation region of the engine body is a low speed and low load operation region. also the in which Ru is the spark plug at a predetermined amount advanced the timing from the point. In this mode, in the case of the low speed and low load operation region, the effective compression ratio is lowered to less than 12 to reliably prevent knocking, and the ignition timing is advanced from the compression top dead center in the same manner as a general engine. By doing so, it is possible to achieve proper combustion at a relatively high compression ratio according to the operating region. In addition, since the effective compression ratio is changed at the intake valve closing timing, it is possible to reduce the pumping loss and improve the fuel consumption. In other words, when the intake valve is slowly closed (or quickly closed) with an engine having a normal compression ratio, the effective compression ratio becomes considerably low, and combustion becomes unstable. For this reason, there are many restrictions on the range in which late closing (or early closing) is possible, and there are restrictions such as insufficient introduction of EGR. However, in the present invention, since the geometric compression ratio is set to be considerably high, even if the effective compression ratio is considerably reduced, the actual compression ratio is still high, so that the combustion stability becomes high. For this reason, the range of the late closing (or early closing) of the intake valve can be widened, and if the valve timing is the same, the EGR rate can be increased as compared with the low compression ratio.

In the aspect in which the geometric compression ratio is set to 13 or more, when the ignition timing is retarded within a predetermined period after the compression top dead center in the high load operation region, the predetermined period includes the low speed and low load operation region. In this case, the ignition timing is preferably set smaller than the advance amount from the compression top dead center. As a result , the retard amount when the ignition timing is retarded in the high load operation region in the low speed region is set to a relatively small value. As a result, in the high load operation region in the low speed region, after shifting to the expansion stroke, it is possible to maintain extremely high torque while avoiding knocking.

  In the aspect in which the geometric compression ratio is set to 13 or more, the low speed range set in the control means is a low speed range when the engine rotation range is divided into three stages of low speed, medium speed, and high speed. The predetermined period is a stroke range of 10% or less after the top dead center has elapsed. In this aspect, the operation region is divided into three stages in the engine speed region, and in the low speed region, the intake valve closing timing is adjusted so that the effective compression ratio is maintained at 12 or more in the throttle fully open region, and the ignition is performed. By retarding the timing within a stroke range of 10% or less after the top dead center has elapsed, it is possible to achieve proper combustion at a relatively high compression ratio according to the operating region. It should be noted that the engine rotation range is not necessarily divided equally.

  In a preferred aspect, the control means switches the ignition timing before the compression top dead center in an engine speed range of medium speed or higher.

  Each aspect includes combustion period shortening means for shortening the combustion period of the air-fuel mixture when the ignition timing is retarded after compression top dead center. In this aspect, the heat generation rate in the expansion stroke is increased as much as possible by the combustion period shortening means, time loss can be suppressed, and high torque can be obtained.

  In a preferred aspect, the combustion period shortening means is turbulent flow generating means for generating turbulent flow in the cylinder. In this aspect, a relatively simple mechanism or control can increase the heat generation rate in the expansion stroke as much as possible, suppress time loss, and obtain a high torque.

  In a preferred embodiment, each cylinder is provided with a plurality of spark plugs, and the combustion period shortening means is a multipoint ignition means for operating a plurality of spark plugs. In this aspect, since the combustion speed can be accelerated by multipoint ignition, the heat generation rate in the expansion stroke can be increased as much as possible, time loss can be suppressed, and high torque can be obtained.

  In a preferred embodiment, a fuel injection valve capable of controlling the injection timing by the control means is provided, and the control means has an operation range of the engine main body at least from a predetermined medium load operation range to a throttle full open range in a low speed range. In the case of the load operation region, split injection is performed in which fuel is injected a plurality of times within a predetermined period from the intake stroke to the compression stroke. In this aspect, by dividing the fuel injection, vaporization and atomization of the fuel injected in the intake stroke is promoted, and a weakly stratified mixture can be formed in the combustion chamber, thereby shortening the combustion time. It is possible to improve the output and fuel consumption.

  In a preferred embodiment, an external EGR system capable of adjusting the amount of external EGR by the control means is provided, and the control means is provided when the operation region of the engine body is a high load operation region including a throttle full open region at least in a low speed region. Introduces external EGR. In this aspect, since the combustion temperature can be lowered by the external EGR, the cooling loss can be reduced while avoiding knocking, and the thermal efficiency is increased. As a result, high torque and fuel consumption can be obtained. That is, when the compression ratio is high, the in-cylinder temperature rapidly rises during the compression stroke, so that knocking is likely to occur. Furthermore, since the heat generated suddenly is absorbed by the wall surface of the cylinder and the like, the heat loss increases. On the other hand, when the burned gas discharged from the exhaust valve is introduced, even when the effective compression ratio is relatively high, the combustion temperature is lowered, and as a result, heat loss is suppressed together with knocking, which is high. Torque and fuel consumption can be maintained.

  In a preferred aspect having the external EGR system, the control means introduces an external EGR at least in a low-speed low-load operation region. In this aspect, coupled with the reduction in the effective compression ratio, it is possible to reduce heat loss as much as possible and maintain high fuel efficiency.

  In a preferred aspect, the control means shifts the closing timing of the intake valve by a predetermined amount from the intake bottom dead center so as to reduce the effective compression ratio at least in the low speed and low load operation region. In this aspect, the effective compression ratio is reduced in the operation region where the combustion state tends to be relatively unstable, and a high expansion ratio is ensured. As a result, it is possible to reduce the pumping loss and improve fuel efficiency while preventing knocking due to the high compression ratio.

In at least a low speed low load operation region, in embodiments of shifting a predetermined amount from the intake bottom dead center the valve closing timing of the intake valve so as to reduce the compression ratio, provided an EGR means for introducing E GR in the cylinder, the EGR means It is preferable that EGR is introduced at least in the low-speed and low-load operation region. In that case, it is possible to reduce the heat loss as much as possible in combination with the decrease in the effective compression ratio and maintain high fuel efficiency. That is, when the compression ratio is high, the in-cylinder temperature rapidly increases during the compression stroke. Here, since the heat generated suddenly is absorbed by the wall surface of the cylinder and the like, the heat loss increases. On the other hand, when the burned gas discharged from the exhaust valve is introduced, the effective compression ratio is lowered, and as a result, the combustion temperature is lowered, so that heat loss is suppressed and high fuel consumption is maintained. It becomes possible to do.

In the aspect in which the EGR means for introducing EGR into the cylinder is provided, the control means preferably sets the air-fuel ratio to the stoichiometric air-fuel ratio in the low-speed low-load operation region. As described above, since the high expansion ratio can be secured by reducing the effective compression ratio under the high compression ratio condition, the fuel consumption can be sufficiently improved even in the operation at the stoichiometric air-fuel ratio. Since a three-way catalyst that is inexpensive and has a high purification rate can be arranged, sufficient exhaust performance can be achieved even in the low-speed and low-load operation region.

In the aspect in which the EGR means for introducing EGR into the cylinder is provided, it is preferable that the low-speed low-load operation region set in the control means includes an idling operation region. In that case, high fuel efficiency can be maintained even in the idling operation region where the frequency of use is high.

  Each aspect includes in-cylinder temperature estimating means for estimating the in-cylinder temperature of the engine body, and the control means sets the intake valve closing timing in the vicinity of the intake bottom dead center at the time of cold start, and performs effective compression. It is preferable that the intake valve closing timing is adjusted and controlled so as to increase the ratio and ensure sufficient intake. In that case, the volumetric efficiency can be increased by increasing the effective compression ratio and securing sufficient intake air, so that sufficient ignition / combustion performance and sufficient torque to increase the engine speed can be obtained. It becomes possible.

  In each embodiment, the piston crown surface of the engine main body is formed on the crown surface peripheral portion and generates a reverse squish flow when moving from the compression top dead center to the expansion stroke, and a central portion of the crown surface Preferably, the control means controls the fuel injection valve so as to inject fuel in the compression stroke. In that case, in the process leading to the compression stroke, a flight space for the injected fuel is secured by the recess formed in the central portion of the crown surface of the piston. A reverse squish flow is formed. As a result, the combustion period is shortened, knocking prevention and heat generation rate in the expansion stroke are increased as much as possible, time loss is suppressed, and high torque and fuel efficiency are improved. Further, in an operating region where the effective compression ratio is 12 or more (11.5 or more for an engine using an octane number of 91 RON or more), the concave portion has a cold flame in the combustion chamber after the compression top dead center has elapsed. It contributes to generation and also becomes an element that further increases output.

  As described above, the present invention is conventionally operated at a high load in a low speed range, which has been used at the expense of output by adopting an expensive mechanism or lowering the effective compression ratio by adjusting the closing timing of the intake valve. In the region, knocking is avoided while maintaining a high compression ratio, so that the fuel efficiency equivalent to that of a diesel engine can be obtained as much as possible while combining low cost and high output. .

[Combustion output mechanism of high compression ratio engine]
First, the relationship between the high compression ratio and the knocking suppression according to the present invention will be described in detail.

  In the course of studying the relationship between knocking and the geometric compression ratio, the inventor is retarded to prevent knocking if the compression ratio is increased so that the ignition timing determined from the knocking limit is after the compression top dead center. We found a phenomenon in which the amount of retarded ignition timing decreased and the output increase due to the compression ratio improvement far surpassed the output decrease due to ignition timing retard to prevent knocking. With respect to this phenomenon, the present inventor has made a hypothesis that the retard amount gradually decreases to a relatively small stroke range when the compression ratio is 13 or more, as indicated by a circle in FIG. FIG. 2 is a graph showing the relationship between the crank angle and the torque (the indicated mean effective pressure (IMEP)) for explaining the hypothesis in the development process of the present invention.

  This hypothesis is that when the ignition timing is retarded after the compression top dead center, the pressure / temperature at the compression top dead center is temporarily increased by increasing the compression ratio. This was based on the idea that the self-ignition is less likely to occur because the piston suddenly drops and the pressure and temperature drop before autoignition occurs.

  In order to verify this hypothesis, the present inventor obtained the graph shown in FIG. 3 as a result of simulating the relationship between the indicated mean effective pressure (IMEP) and the ignition timing by numerical simulation. FIG. 3 is a graph showing a simulation result showing the relationship between the ignition timing and IMEP.

  As shown in FIG. 3, when comparing compression ratios 11 and 12, when comparing 12 and 13, IMEP slightly increases, whereas when comparing 13 and 14, IMEP is As a result of numerical simulation, it has been clarified that the increase ratio of IMEP decreases when compared with 14 and 15 when compared with 14 and 15. In order to verify this output change, the present inventor examined the heat generation rate at each compression ratio.

  FIG. 4 is a graph showing the relationship between the heat generation rate and the crank angle when an engine having a compression ratio of 11, 13, 14, 15 is ignited at 8 ° CA after the compression top dead center has elapsed.

  As shown in FIG. 4, when the compression ratio is 11 or 13, the heat generation rate from the compression top dead center to the ignition timing is gradually increasing, whereas when the compression ratio is 14, immediately before the ignition timing. The heat generation rate of is significantly increased. From this result, the compression ratio is set higher from a certain value (ε = 13 in the case of 96 RON) to a certain value (ε = 14 in the case of 96 RON), thereby exceeding the surrounding cooling loss due to the pressure increase due to the piston rising. It can be seen that a cold flame reaction with a slight exothermic reaction occurs.

  From the results of FIGS. 3 and 4, when the compression ratio is high and the ignition timing is retarded after the compression top dead center, the combustion process after the compression top dead center becomes multistage ignition, and in particular, the predetermined compression ratio. In the case of an octane number of 96 RON and a geometric compression ratio of 14, for example, it has been revealed that the cold flame reaction becomes significant. The effect of suppressing the deterioration of knock of the cold flame reaction will be described below.

  FIG. 5 is a graph simulating the combustion process after the compression top dead center has elapsed at a high compression ratio. The upper graph shows the relationship between pressure and time, and the lower graph shows the relationship between the number of moles increase and time. The calculated values are obtained by preparing a high-temperature and high-pressure constant capacity vessel and calculating changes in pressure and moles over time.

As shown in FIG. 5, when the elapsed time reaches t1 after the piston passes through the compression top dead center, a cold flame reaction occurs and the pressure slightly increases. In the time when this cold flame reaction occurs, the volume is constant and the number of moles increases.
PV = nRT (1-1)
However, as apparent from P: pressure, V: volume, n: number of moles, R: gas constant, T: temperature, the temperature does not increase as much as the pressure increases. For this reason, in relation to the temperature, there is no temperature increase that increases the pressure even in the end gas in the cylinder, and self-ignition is less likely to occur. In the combustion chamber (constant capacity device), after a predetermined time elapses (t2), a hot flame reaction occurs due to a chain reaction, resulting in a multistage ignition phenomenon in which the pressure rapidly increases.

  Next, when the piston reaches compression top dead center, the in-cylinder temperature changes as shown in FIG. FIG. 6 is a contour line showing the temperature distribution of the combustion chamber when the compression top dead center is reached.

  As shown in FIG. 6, the combustion chamber when the piston reaches compression top dead center becomes hot at the center due to the cold flame reaction, but the peripheral portion (end gas portion) is affected by the wall temperature. Since the cold flame reaction does not easily proceed, the in-cylinder temperature in the peripheral portion remains at about 800K. For this reason, in the process in which the cold flame reaction is occurring, the in-cylinder temperature in the peripheral portion remains relatively low, so that the combustion proceeds and the knocking deterioration is suppressed.

  Next, while the cold flame reaction proceeds in the combustion chamber, formaldehyde (HCHO) is generated. Since this formaldehyde absorbs OH radicals that cause knocking when the temperature of the combustion chamber is 900K or less, knocking is suppressed.

  FIG. 7 is a graph showing the in-cylinder pressure during combustion and the adiabatic compression temperature history of the end gas portion in the peripheral portion. The upper row shows the relationship between the pressure and the crank angle, and the lower row shows the relationship between the end gas temperature and the crank angle. Show.

  As shown in FIG. 7, in the process of the piston of a cylinder from the bottom dead center through the compression top dead center to the bottom dead center, the pressure increases by a predetermined crank angle from the compression top dead center, and the temperature also increases accordingly. The temperature rises at the same timing, but the temperature of the end gas portion of the combustion chamber does not exceed 900K unless the intake air temperature is extremely high. Therefore, even in an engine with a high compression ratio such that the ignition timing determined from the knocking limit is after the compression top dead center, a multi-stage ignition phenomenon occurs, so that formaldehyde generated in the cold flame reaction contributes to suppression of knocking. all right.

As described above, in an engine with a high compression ratio such that the ignition timing determined from the knocking limit is after compression top dead center, as a knocking suppression mechanism,
(1) Due to the cold flame reaction, the combustion chamber does not rise as much as the pressure rises.
(2) Since the cold flame reaction mainly occurs in the center of the combustion chamber, the temperature of the end gas portion is relatively low,
(3) Since the combustion chamber remained at a predetermined temperature (900K) or less even after the piston passed through compression top dead center, it was found that formaldehyde consumed OH radicals. Therefore, the present inventors calculated the knocking limit by adding these knocking suppression mechanisms to the calculation based on the conventional chemical reaction.

  The circles in FIG. 3 indicate the knocking limit simulation results at each compression ratio. As indicated by the circles in FIG. 3, the knock limit at each compression ratio 11 to 15 is such that when the compression ratios 11 and 12 are compared, and when 12 and 13 are compared, the retard amount is substantially the same. On the other hand, when 13 and 14 were compared, it was found that the retard amount hardly changed. Furthermore, when comparing 14 and 15, it was found that the retard amount increased again.

  From these results, it was found that the amount of retard for suppressing knocking was dependent on the amount of heat generated in the cold flame reaction, and gradually decreased at a certain compression ratio and increased again when the compression ratio was exceeded. I found out that

  Next, the present inventor performed a simulation on the relationship between the cold flame reaction and the torque.

  FIG. 8 is a graph showing the relationship between the heat generation rate and the crank angle when the compression ratio is 14, and FIG. 9 is a PV diagram when the compression ratio is 14 based on numerical simulation. In each figure, C11 shows a case where a cold flame reaction is caused after the elapse of the compression top dead center as in the case of an actual engine, and C12 shows a case where a cold flame reaction is not intentionally caused.

  As shown in FIG. 8, when a cold flame reaction is caused by the ignition retard after the compression top dead center has elapsed, the heat release rate gradually increases immediately after the compression top dead center has elapsed, and after ignition (compression top dead center). After the elapse of 8 ° CA) Ascending without pre-ignition.

  As a result of calculating the PV characteristics based on this assumption, as shown in FIG. 9, the PV characteristics are combusted while the pressure after the compression top dead center is high, and compared with the case where no cold flame reaction occurs. It was found that time loss was reduced.

  From these simulation results, when the compression ratio is set high and the ignition timing is retarded after the compression top dead center, when the compression ratio is 14, the increase in the heat generation rate due to the cold flame reaction is increased and the time loss is reduced. It has been found that a high torque can be obtained with a reduction.

  Next, the inventor of the present invention examined how the relationship between the compression ratio and the knocking limit as described above changes depending on the octane number.

  FIG. 10 is a graph showing the relationship between the compression ratio and the amount of heat generated by the cold flame reaction for each octane number.

  Referring to FIG. 10, as a result of measuring the amount of heat by the cold flame reaction by combining the octane number and the compression ratio, when a fuel having an octane number of 96 RON is used, the cold flame reaction is caused by an engine having a compression ratio of 12.5 or more. It became prominent, and the cold flame reaction gradually decreased in an engine having a compression ratio of 15 or more. Based on this measured value, when 91 RON and 100 RON are calculated, when a fuel with an octane number of 91 RON is used, the cold flame reaction becomes significant in an engine with a compression ratio of 12.0 or more, and the compression ratio is When an engine with a compression ratio of 14.5 or higher is used and fuel with an octane number of 100 RON is used, the cold flame reaction becomes remarkable with an engine with a compression ratio of 13.0 or higher, and the compression ratio is 15.5. The cold flame reaction gradually decreases with the above engine. Based on the graph of FIG. 10, the output of the knocking occurrence point was calculated for each octane number.

  FIG. 11 is a graph showing the relationship between the compression ratio calculated based on the graph of FIG. 10 and the indicated mean effective pressure (IMEP) for each octane number.

  Referring to FIG. 11, when a fuel with an octane number of 96 RON is used, a cold flame reaction becomes significant in an engine having a compression ratio of 13 or more and 15 or less. The output is improved by the effect and the improvement of the compression ratio. Similarly, when a fuel with an octane number of 100 RON is used, an engine with a compression ratio of 13.5 to 15 and when a fuel with an octane number of 91 RON is used, the compression ratio is 12.5 to 13.5 or less. The engine improves the compression ratio output just before the cold flame reaction occurs.

  In particular, it has been verified that the output is improved in the vicinity of a compression ratio of 14 at which the cold flame reaction is most prominent in fuels having an octane number of 96 RON and 100 RON.

  Next, the upper limit of the compression ratio will be described.

  At a compression ratio where the ignition retard is after compression top dead center, the output increases due to the cold flame reaction, but pre-ignition is likely to occur when the temperature and pressure are high and the time is long. For example, if the engine is temporarily stopped in the parking area when it is warm and restarted when the intake air temperature is rising, the intake air temperature may become abnormally high. The temperature may rise rapidly and pre-ignition may occur. Recently, an engine having a variable valve timing system (VVT) capable of adjusting the closing timing of the intake valve has become widespread. However, when the throttle is fully open at low speed, the intake valve closes at the bottom of the intake bottom dead center. Since it is 30 ° CA or less later, the difference between the effective compression ratio and the geometric compression ratio is 1 or less as shown in FIG. For this reason, even if the method of changing the closing timing of the intake valve is adopted, the range in which the effective compression ratio can be reduced is limited, and the upper limit of the geometric compression ratio is set based on some standard. It is preferable. In addition, there is a low-speed and high-load operation region where the effective compression ratio cannot be reduced depending on the operation state. Therefore, in the present invention, the combination of the geometric compression ratio, the effective compression ratio, and the octane number is set as shown in Table 1, thereby achieving both improvement in output and suppression of knocking.

  Next, the influence when a supercharger is attached to the engine will be described.

  FIG. 13 is a graph showing the distribution of the ignition delay τ in relation to pressure and temperature.

  First, Arrhenius function

(However, τ: ignition delay of mixture at instantaneous temperature and pressure, A, n: mixture constant specific to reaction conditions, p: absolute pressure, E: apparent activation energy, R: gas constant, T: temperature)
Based on the above, the constant A in the formula (1-2) is set to a constant proportional to the octane number, and the ignition delay τ in the case of using 96 RON fuel is obtained, and τ is set to 19 steps Rτ1 to Rτ19 every 10 milliseconds, (MPa) vs. temperature (Kelvin). Rτ indicates that the larger the subscript, the more likely self-ignition occurs.

  Next, we investigated the pressure and temperature at the compression top dead center of a normal engine without a supercharger and a turbocharged engine equipped with a turbocharger.

  As a normal engine, when the geometric compression ratio is 13, the pressure and temperature Ap1, Ap2 at the compression top dead center of 14 were obtained.

  As for the engine with a supercharger, when the geometric compression ratio is 12, the case of 13 and the case of 14 are calculated for the cases where the supercharging pressure is 120 KPa and 140 KPa, respectively. Temperatures Ap3 to Ap8 were determined.

  As shown in FIG. 13, the self-ignition points Ap1 (ε = 13) and Ap2 (ε = 14) of the normal engine are located on the lines of Rτ4 and Rτ5 of the ignition delay τ, respectively. On the other hand, the self-ignition points Ap3 (ε = 13) and Ap4 (ε = 14) of the supercharged engine having the same geometric compression ratio are on (or in the vicinity of) the Rτ5 and Rτ6 lines of the ignition delay τ, respectively. I found that it is located at. That is, in the supercharged engine, the ignition delay at the compression top dead center is shortened by about 10 milliseconds as compared with a normal engine having the same geometric compression ratio, and self-ignition tends to occur. This tendency is due to the pressure and temperature Ap3, Ap5 at the compression top dead center of a turbocharged engine with a geometric compression ratio of 13, and the compression top dead center of a turbocharged engine with a geometric compression ratio of 12. It was more remarkable when the pressure at the temperature and the temperature Ap7, Ap8 were compared.

  From the above results, the self-ignition characteristics of the high compression ratio engine can be increased by reducing the geometric compression ratio and the effective compression ratio by 1 compared to the normal engine. It was verified that the engine can be as good as a normal engine.

  In addition, since the constant A in the equation (1-2) is a value proportional to the octane number, even when the geometric compression ratio or the effective compression ratio is changed due to the difference in the octane number, the predetermined value is decreased from those values. It has been verified that similar results can be obtained.

  In recent years, biofuels have been developed that make methyl esters from ethanol (ethyl alcohol), methanol (methyl alcohol), edible oil, etc., and use them as fuel for automobiles. Since the octane number tends to increase, the technical idea of the present invention can be applied.

Based on the above findings, the following embodiment has been completed.
[Embodiment]
Hereinafter, preferred embodiments of the present invention will be described with reference to the accompanying drawings.

  FIG. 14 is a configuration diagram showing a schematic configuration of a four-cycle spark ignition gasoline engine 10 according to an embodiment of the present invention, and FIG. 15 is a cross-sectional view showing a structure of one cylinder of the engine body 20 according to FIG. It is a schematic diagram.

  Referring to FIGS. 14 and 15, the illustrated spark ignition gasoline engine 10 includes an engine body 20 and a control unit 100 for controlling the engine body 20.

  The engine body 20 integrally includes a cylinder block 22 that rotatably supports the crankshaft 21 and a cylinder head 23 that is disposed above the cylinder block 22. Are provided with a plurality of cylinders 24.

  Each cylinder 24 is provided with a piston 26 connected to the crankshaft 21 via a connecting rod 25 and a combustion chamber 27 formed in the cylinder 24 by the piston 26. In the present embodiment, the geometric compression ratio of each cylinder 24 is set to 13.

  Referring to FIG. 15, the engine body 20 according to the present embodiment has a cylinder bore center Z (see FIG. 16) of the cylinder 24 as viewed from the side in which the rotation direction of the crankshaft 21 is clockwise (that is, the state of FIG. 15). ) Is offset to the right from the rotation center O of the crankshaft 21. This offset amount S is set to, for example, 1 mm to 2 mm when the bore diameter of the cylinder 24 is 70 mm.

  FIG. 16 is a schematic plan view showing the cylinder 24 in an enlarged manner.

  Referring to FIG. 16, a ceiling portion of combustion chamber 27 is configured for each cylinder 24 on the lower surface of cylinder head 23, and this ceiling portion has two inclined surfaces extending from the central portion to the lower end of cylinder head 23. It is a pent roof type.

  A fuel injection valve 32 that receives a fuel injection pulse from the control unit 100 and injects fuel corresponding to the pulse width into the combustion chamber 27 is provided at the side of the combustion chamber 27.

  Each cylinder 24 is provided with three spark plugs 34 that are fixed to the cylinder head 23 and emit a spark in the combustion chamber 27. The spark plugs 34 are arranged along the cylinder diameter parallel to the ridge line portion of the piston 26, the center one is disposed on the cylinder bore center Z, and both sides are disposed on the side edge of the combustion chamber 27. Each ignition plug 34 is connected to an ignition circuit 35 (see FIG. 15) that can control the ignition timing by electronic control. The ignition circuit 35 is controlled by the control unit 100, whereby the ignition plug 34 is The ignition control is selectively performed.

  FIG. 17 is an explanatory diagram showing the airflow in the combustion chamber 27 according to this embodiment, where (A) shows the initial stage of the compression stroke and (B) shows the initial stage of the expansion stroke.

  Referring to FIGS. 16 and 17, one inclined surface (the inclined surface on the right side in FIGS. 17A and 17B) 27a constituting the ceiling portion of the combustion chamber 27 is a set of two independent ones. The intake port 28 is opened, and two exhaust ports 29 are opened on the other inclined surface (left inclined surface in FIGS. 17A and 17B) 27b. Are provided with an intake valve 30 and an exhaust valve 31. The intake port 28 is a straight port that extends linearly from the combustion chamber 27 to the upper right in FIG. 17, and is configured to be away from the cylinder bore center Z toward the intake upstream side in the cross section shown in FIG.

  On the crown surface of the piston 26, there is provided a squish area constituting surface that is inclined along the inclined surface of the cylinder head 23 within a predetermined range of the peripheral portion on the intake side and a predetermined range of the peripheral portion on the exhaust side. Further, a raised portion 33 is provided inside the squish area constituting surface.

  The raised portion 33 is provided in a predetermined range including the projection surfaces of the intake valve 30 and the exhaust valve 31. The skirt portion of the raised portion 33 is formed in a mountain shape having a pair of inclined surfaces 33a and 33b substantially parallel to both inclined surfaces 27a and 27b of the combustion chamber ceiling. Both inclined surfaces 33a and 33b are raised in a pent roof shape corresponding to the ceiling surface of the combustion chamber 27, and have a shape having a top portion 33c at a portion corresponding to the ridgeline of the ceiling surface. And the recessed part 264 is formed over the top part 33c of this piston crown surface, and the inclined surfaces 33a and 33b of the both sides. The bottom surface of the recess 264 is a curved surface that approximates a spherical surface. And although the planar shape of the recessed part 264 is a shape close | similar to a circle | round | yen, it is an ellipse whose diameter of the direction orthogonal to this is a little longer than the diameter of the direction along the top part (ridgeline) 33c.

  For this reason, in this embodiment, there exist the following effects.

  That is, the intake air sucked into the combustion chamber 27 by the lowering of the piston 26 in the intake stroke forms two types of flows as indicated by Ta1 and Ta2 in FIG. One flow Ta1 flows mainly into the combustion chamber 27 from the side near the spark plug 34 of the intake port 28 opening, flows toward the combustion chamber peripheral wall surface on the exhaust side, and then proceeds downward along the exhaust side peripheral wall surface. After that, it flows along the piston crown surface to the intake side, and from there upwards. The other flow Ta2 is a weak flow that has flowed out from the cylinder bore peripheral side of the throat portion. With these flows Ta1 and Ta2, as shown in FIG. 17B, a forward tumble flow Ta1 swirling counterclockwise and a reverse tumble flow Ta2 swirling clockwise are generated. Yes.

  Next, when moving to the compression stroke, as the piston 26 rises, the forward and reverse tumble flows Ta1 and Ta2 are swung in the combustion chamber 27 in the vertical direction, and the downstream sides of the pistons 26 turn in the opposite direction from the center side of the cylinder bore.

  The forward tumble flow Ta1 is larger and stronger than the reverse tumble flow Ta2 in the initial to middle stage of the compression stroke. However, as the compression stroke proceeds and the piston 26 approaches the top of the combustion chamber, the forward tumble flow Ta1 The center gradually becomes smaller as it moves to the exhaust side. At the end of the compression stroke or the early stage of the expansion stroke where the piston 26 is near the top dead center, the forward and reverse tumble flows Ta1 and Ta2 have the same magnitude and strength, and the exhaust side and the intake side in the combustion chamber 27 It will be in the state which turns into a mutually reverse direction. By these forward and reverse tumble flows Ta1 and Ta2, at the end of the compression stroke in which the piston 26 approaches the top dead center, the center of the combustion chamber 27 from the squish area between the inclined surfaces 27a and 27b of the combustion chamber ceiling and the piston crown surface. In the initial stage of the expansion stroke in which the forward squish flow is generated in the direction toward the side (the direction opposite to the white arrows Ra and Rb in FIG. 16B) and the piston 26 starts to descend after reaching the top dead center, Reverse squish flows Ra and Rb from the center of the combustion chamber 27 toward the squish area are generated as indicated by white arrows in FIG.

  In this case, the forward and reverse tumble flows Ta1 and Ta2 flow in the opposite direction to the forward squish flow and in the same direction as the reverse squish flow Ra and Rb. Therefore, the forward squish flow at the end of the compression stroke is weakened and the reverse squish flow is reversed. The effect of strengthening the reverse squish flow Ra, Rb is exhibited.

  By strengthening the reverse squish flows Ra and Rb in this manner, the combustion speed in the squish area is sufficiently increased, the main combustion speed of the flame is increased, and rapid combustion is realized. Moreover, since the normal tumble flow Ta1 is moderately weakened, the initial combustion rate is not so high, and the spontaneous combustion of the air-fuel mixture in the end gas zone is not induced. That is, the initial combustion period is not so shortened, and the main combustion period is significantly shortened, so that knocking is suppressed in advance, time loss is reduced by rapid combustion, and thermal efficiency is improved.

  Further, since the pair of inclined surfaces 33a and 33b are parallel to the inclined surfaces 27a and 27b of the combustion chamber ceiling portion, flame propagation is performed uniformly in the space between them, which is effective in preventing detonation.

In addition, since the concave portion 264 is formed in the central portion of the crown surface of the piston 26 to secure the flight space of the injected fuel, the mixture is disturbed in the cylinder when the fuel is injected, Further, the combustion period can be shortened. Further, in the operation region where the effective compression ratio ε _r is operated at 12 or more, the concave portion 264 contributes to the generation of the cool flame in the combustion chamber 27 after the compression top dead center has elapsed, and is also an element that further increases the output. Become.

  Next, each intake valve 30 is configured to be driven by a valve operating mechanism 40. The valve operating mechanism 40 has a variable camshaft timing mechanism (VCT) 42 that can change the opening / closing timing of the intake valve 30 steplessly, and a VVE (Variable) that can change the lift amount (opening amount) of the intake valve 30 steplessly. Valve Event) 43.

  FIG. 18 is a perspective view showing a specific configuration of the valve mechanism 40 according to the embodiment of FIG.

  With reference to the figure, the valve operating mechanism 40 includes a cam shaft 41a extending along the direction in which the cylinders 24 are arranged (see FIG. 1), and a VCT 42 and a VVE 43 are incorporated in the cam shaft 41a. .

  The VCT 42 includes a rotor (input member) 42a fixed to the end of the camshaft 41a, a casing (output member) 42b disposed concentrically on the outer periphery of the rotor 42a, and the casing 42b. And a sprocket 42c disposed on the outer periphery of the sprocket 42 relatively rotatably. A chain 42d for transmitting a driving force from the crankshaft 21 (see FIG. 14) is wound around the sprocket 42c. Further, a hydraulic oil chamber (not shown) is formed between the rotor 42a and the casing 42b, and the rotor 42a and the casing 42b are integrally rotated or relatively rotated by hydraulic control of the electromagnetic valve 42e. It can be switched to operation. Thus, the VCT 42 constitutes an operation timing variable mechanism that can simultaneously change the valve opening start timing and the valve closing timing of the intake valve 30. As will be described later, the electromagnetic valve 42e is driven and controlled by the control unit 100, and the rotor 42a and the casing 42b are connected / disconnected by this driving control.

  Next, the VVE 43 includes a pair of intake cams 43 a and 43 b provided on each intake valve 30. Each intake cam 43a is fixed to the camshaft 41a. The other intake cam 43b is attached to the camshaft 41a via a cam journal 43c so as to be relatively rotatable.

  FIG. 19 is a cross-sectional view showing the main part of the VVE in FIG. 18, (A) shows when the lift amount is 0 in the large lift control state, and (B) shows the maximum lift amount in the large lift control state. (C) shows the time when the lift amount is 0 in the small lift control state, and (D) shows the time when the lift amount is maximum in the small lift control state.

  Referring to FIG. 18 and FIGS. 19A to 19D, in order to synchronize the intake cam 43b attached to the camshaft 41a so as to be rotatable relative to the camshaft 41a, An eccentric cam 43d provided for each cylinder 24 is fixed. As is apparent from FIGS. 19A to 19D, the eccentric cam 43d is eccentric with respect to the cam shaft 41a. An offset link 43e is rotatably attached to the outer periphery of the eccentric cam 43d. A protrusion 43f protruding in the radial direction is integrally provided on the outer peripheral portion of the offset link 43e. A connecting pin 43g parallel to the camshaft 41a passes through the projecting portion 43f, and one end of each of the link arms 43h and 43i is rotatable on both side surfaces of the offset link 43e by the connecting pin 43g. Is attached. One link arm 43h connects the offset link 43e and the intake cam 43b, and the other end of the link arm 43h is rotatable to the vicinity of the bulging portion of the intake cam 43b by a pin 43j parallel to the cam shaft 41a. It is connected to. The other link arm 43i is for connecting the offset link 43e to the eccentric shaft 43k that changes the phase of the offset link 43e. With respect to the end of the control arm 43m fixed to the eccentric shaft 43k, The other end is rotatably connected by a pin 43n parallel to the camshaft 41a.

  As shown in FIG. 18, a fan-shaped worm wheel 43p is fixed in the middle of the eccentric shaft 43k, and a worm gear 43q meshing with the worm wheel 43p is driven to rotate by a stepping motor 43r. Yes. As will be described later, the stepping motor 43r is driven and controlled by the control unit 100. By this driving control, the phase of the control arm 43m is determined, and thereby the phase of the offset link 43e is determined. Therefore, the rotational trajectory of the intake cam 43b that drives the tappet 36 changes in the axial direction of the intake valve 30, and the valve lift amount is changed steplessly.

  Referring to FIG. 19B, the tappet 36 provided on the valve stem 30 a of each intake valve 30 is fixed to the end of the valve stem 30 a of the intake valve 30. On the other hand, the valve stem 30a of the intake valve 30 is guided by a known valve guide 30b. A spring seat portion 30c is integrally formed on the outer periphery of the valve guide 30b. The spring seat portion 30c is contracted between the spring seat portion 36a formed in the inner back portion of the tappet 36. A valve spring 30d is seated.

  The intake cam 43b is joined to the tappet 36 and receives the urging force of the valve spring 30d.

  In this state, as shown in FIGS. 19A and 19B, when the eccentric shaft 43k and the control arm 43m are rotated by the stepping motor 43r and the pin 43n is positioned below the eccentric shaft 43k, the intake cam 43b The swing angle becomes large, and a large lift control state in which the lift amount of the valve at the lift peak is the largest is achieved. Further, when the pin 43n is moved upward by the rotation of the control arm 43m or the like, the swing angle of the intake cam 43b is reduced accordingly, and as shown in FIGS. 19C and 19D, the pin 43n Is positioned above the camshaft 41a, the small lift control state with the smallest valve lift amount is achieved.

  In the large lift control state shown in FIGS. 19A and 19B, the intake cam 43b presses the tappet 36 on the tip side of the cam nose, as shown in FIG. 19B, and the intake valve 30 passes through the tappet 36. The lift amount of the intake valve 30 becomes 0 as shown in FIG. 4D, when the lift valve is lifted greatly (the intake cam 43b lifts the intake valve 30 through the tappet 36). Swings between states. Similarly, in the case of FIGS. 19C and 19D, which are in the small lift control state, swinging occurs between the lift peak state (the tappet 36 is pressed on the base end side of the cam nose) and the lift amount 0 state (same as above). (See Figures (C) and (D)).

  20 schematically shows the control states of FIGS. 19B and 19D, where FIG. 20A corresponds to the large lift control position, and FIG. 20B corresponds to the small lift control position. 20A and 20B, the control arm 43m, the connecting link 43h, and the link arm 43i are simply represented by straight lines, and the center of the eccentric cam 43d (the center of the outer ring of the offset link 43e) is rotated. The trajectory is shown as T0.

  First, the profile of the intake cam 43b itself will be described with reference to FIG. 20A. On the peripheral surface of the intake cam 43b, a base circle surface (base circle section) θ1 whose curvature radius is constant within a predetermined angle range, Following the θ1, a cam surface (lift section) θ2 having a gradually increasing radius of curvature is formed.

  The solid line in FIG. 20A shows the state of FIG. 19B in which the intake valve 30 is in the vicinity of the lift peak. At this time, the pin 43j is pulled up most by the connecting link 43h, and the intake cam 43b The cam nose tip side of the cam surface θ2 is in contact with the tappet 36. On the other hand, the phantom line shows a state where the valve lift amount H is 0 (FIG. 19A). At this time, the base circle surface θ1 of the intake cam 43b is in contact with the tappet 36 and the intake valve 30 is closed. It is in a state.

  When the camshaft 41a (eccentric cam 43d) rotates in the clockwise direction in the figure, one end side (lower end side in the figure) of the offset link 43e moves along the axis X of the camshaft 41a as indicated by the arrow in the figure. Although revolving around, the displacement of the other end of the offset link 43e is regulated by a link arm 43i connected thereto. That is, the link arm 43i swings between the position of the solid line and the position of the phantom line around the pin 43n positioned below the eccentric shaft 43k, and accordingly, the other end side of the offset link 43e. The (connecting pin 43g) reciprocates around the pin 43n every time the eccentric cam 43d rotates once (the movement locus of the connecting pin 43g is indicated as T1).

  In accordance with the reciprocating arc motion T1 of the connecting pin 43g, the other end portion (pin 43j) of the connecting link 43h whose one end portion is connected to the offset link 43e by the same connecting pin 43g reciprocates along a locus indicated by T2 in the drawing. The intake cam 43b, which moves in an arc and is connected to the connection link 43h by the pin 43j, swings between the position of the solid line and the position of the phantom line in the figure. That is, when the connecting pin 43g moves upward, the pin 43j is pulled upward by the connecting link 43h, and the cam nose of the intake cam 43b pushes down the tappet 36, thereby causing the valve spring 30d (see FIG. 19B). The intake valve 30 is lifted while compressing.

  On the other hand, when the connecting pin 43g moves downward, the pin 43j is pushed downward by the connecting link 43h, and the cam nose of the intake cam 43b rises. Therefore, due to the reaction force of the compressed valve spring 30d, The tappet 36 is pushed up and moves upward so as to follow the rise of the cam nose, the intake valve 30 is pulled up, and the intake port 28 is closed.

  That is, in the large lift control state, the intake cam 43b swings greatly so as to press the tappet 36 by substantially the entire base circle surface θ1 and cam surface θ2 of the peripheral surface, and thus corresponds to such a large swing angle. This increases the lift amount of the valve.

  Further, from the above-mentioned large lift control state, the control arm 43m is rotated upward about the axis of the eccentric shaft 43k until it becomes substantially horizontal, as shown in FIGS. 19D and 20B. When the pin 43n, which is the rotation axis of the link arm 43i, is positioned closer to the front side in the rotational direction of the camshaft 41a than the large lift control state, the small lift control state is established. In FIG. 20B, as in FIG. 20A, the state where the intake valve 30 is in the vicinity of the lift peak is indicated by a solid line, and the state where the lift amount H is 0 is indicated by a virtual line.

  In FIG. 20B, when the camshaft 41a (eccentric cam 43d) rotates, the displacement of the connecting pin 43g of the offset link 43e is restricted by the link arm 43i, as in the large lift control state, and the side of the eccentric shaft 43k. A reciprocating arc motion T3 is performed around the pin 43n positioned at (the link arm 43i reciprocates between the solid line position and the virtual line position in the figure). Along with the reciprocating arc motion T3 of the connecting pin 43g, the pin 43j of the connecting link 43h performs the reciprocating arc motion T4, and the intake cam 43b connected to the connecting link 43h by the pin 43j is positioned in the solid line in the figure. The intake valve 30 is opened and closed by swinging between the imaginary line and the position of the imaginary line.

  That is, in the small lift control state, the swing angle of the intake cam 43b is smaller than that in the large lift control state, and the intake cam 43b has a base circle surface θ1 on its peripheral surface and a cam surface θ2 continuous therewith. The tappet 36 is pressed only by a part, and the lift amount of the valve is reduced.

  The valve mechanism 40 as described above is also provided in the exhaust valve 31, and the closed timing of the exhaust valve 31 is advanced from the exhaust top dead center so that the burned gas in the cylinder remains in the intake stroke transition. It is possible to make it an internal EGR.

  Next, referring to FIGS. 14 and 15, the branch intake pipe 133 of the intake manifold 132 is connected to the intake port 28 of the engine body 20. The branch intake pipe 133 is provided for each cylinder 24, and each branch intake pipe 133 is connected to the intake manifold 132 in a state where an equal-length intake path is formed. In the illustrated embodiment, the downstream end of the branched intake pipe 133 is bifurcated so as to correspond to the intake port 28 of each cylinder 24 constituted by two. An open / close valve 134 is provided at the upstream side merge portion of the branch intake pipe 133. The on-off valve 134 is embodied by a three-way solenoid valve, and is configured such that the actuator 135 can individually open and close the aggregate portion of the branch intake pipe 133 by a desired amount. On the other hand, as shown in FIG. 15, a well-known swirl on / off valve 133a is provided at one branch portion of the bifurcated branch intake pipe 133. The swirl on / off valve 133a is driven to open and close by being driven by an actuator 133b. When one of the branch portions of the branch intake pipe 133 is closed by the swirl on / off valve 133a, the swirl on / off valve 133a passes through the other branch. The swirl is generated in the combustion chamber 27 by the intake air, and the swirl is weakened as the swirl generating on-off valve 133a is opened.

  An intake passage 136 for introducing fresh air into the intake manifold 132 is connected to the upstream side of the intake manifold 132. A throttle valve 137 is provided in the intake passage 136.

  The exhaust port 29 is connected to a bifurcated branch exhaust pipe 151 formed in pairs for each cylinder 24. The downstream end of each branch exhaust pipe 151 is connected to the exhaust manifold 152. An exhaust passage 153 for discharging burned gas is connected to the exhaust manifold 152.

  Next, an external EGR system 160 for returning the exhausted burned gas to the intake manifold 132 is provided between the intake manifold 132 and the exhaust manifold 152.

  The external EGR system 160 is connected to a return passage 161 formed between the intake manifold 132 and the exhaust manifold 152, and includes an EGR cooler 162, an EGR valve 163, and an actuator 164 that drives the EGR valve 163. Valve system.

  Next, referring to FIG. 14, the engine 10 according to the present embodiment is provided with a turbocharger 180 as a supercharger.

  The turbocharger 180 includes a turbine section 181 driven by the pressure of exhaust exhausted from the exhaust manifold 152, and a compressor section 182 that is driven by the turbine section 181 and derives supercharged fresh air to the intake manifold 132. The intercooler 183 that cools fresh air is provided between the compressor section 182 and the intake manifold 132. Basically, conventional turbochargers and intercoolers are applied as they are. Is possible.

  Referring to FIG. 14, in order to detect the operating state of engine body 20, an air flow sensor SW <b> 1 is provided in intake passage 136, and an intake air temperature sensor for predicting the in-cylinder temperature downstream of on-off valve 134. SW2 (see FIG. 15) is provided. The cylinder block 22 is provided with a crank angle sensor SW3 for detecting the rotation speed of the crankshaft 21 and an engine water temperature sensor SW4 for detecting the temperature of the cooling water (see FIG. 15). Further, the exhaust passage 153 is provided with an oxygen concentration sensor SW5 for controlling the air-fuel ratio.

  The engine body 20 is provided with a control unit 100 as control means. The control unit 100 is connected with an air flow sensor SW1, an intake air temperature sensor SW2, a crank angle sensor SW3, an engine water temperature sensor SW4, an oxygen concentration sensor SW5, and an accelerator opening sensor SW6 for detecting engine load as input elements. ing. Each of these sensors SW1 to SW6 is a specific example of the driving state detection sensor in the present embodiment. On the other hand, the control unit 100 is connected with a valve operating mechanism 40, an actuator 133b of the swirl generating on-off valve 133a, an actuator 135 of the on-off valve 134, an actuator of the throttle valve 137, and an actuator 164 of the external EGR system 160 as control elements. Yes.

  Referring to FIG. 14, the control unit 100 includes a CPU 101, a memory 102, an interface 103, and a bus 104 that connects these units 101 to 103, and an operation state is determined by a program and data stored in the memory 102. The operation state determination means for determining is functionally configured.

  The memory 102 stores various control maps, which will be described in detail later. Based on these stored maps, the engine body 20 is suitably operated according to the operating state.

  FIG. 21 is a graph showing the relationship between the engine speed N that is the basis of the control map and the required torque in the embodiment of FIG.

Referring to FIG. 21, in the illustrated embodiment, an ignition retard operation region (low-speed medium and high load operation region) A in which the ignition timing is retarded after compression top dead center, and a normal ignition operation region B in which ignition is performed before compression top dead center. And is roughly divided. The ignition retard operation region A is divided into three stages of a low speed region, a medium speed region, and a high speed region when the engine speed N is divided into three stages, that is, a predetermined medium and high load operation region A1 to a throttle fully open region A _WOT. Is set in the range. In the present embodiment, as will be described later, in this ignition retard operation region A, the effective compression ratio ε _r obtained at the intake valve closing timing defined by the valve lift amount of 1 mm is determined by the closing timing of the intake valve 30. It is designed to be operated while maintaining 12 or more.

  On the other hand, the normal ignition operation region B has a low speed and low load operation region B1 as indicated by a broken line, and the low speed and low load operation region B1 includes an idling operation region B2.

FIG. 22 is a timing chart showing a control example of the effective compression ratio ε _r that is the basis of the control map in the embodiment of FIG.

  With reference to FIG. 22, in this embodiment, the valve operating mechanism 40 provided with VVE43 is employ | adopted as mentioned above. By using the valve operating mechanism 40, the valve opening timing and the valve lift amount of the intake valve 30 are controlled steplessly.

In the present embodiment, in the ignition retard operation region A, as is apparent from FIG. 22, the closing timing of the intake valve 30 is retarded until just after the intake bottom dead center, thereby maintaining the effective compression ratio ε _r at 12 or more. On the other hand, as shown in FIGS. 21 and 23, the ignition timing is retarded after the compression top dead center, so that knocking in the ignition retard operation region A is surely prevented.

On the other hand, in the remaining operation region B, in principle, the intake valve 30 is closed early, and the effective compression ratio ε _r is lowered to less than 12 (for example, about 8). As a result, the pumping loss can be reduced. Here, if the effective compression ratio is ε _r ,

(2) where
ε : Geometric compression ratio v _s : Stroke volume (m ³ )
v _c : Clearance volume (m ³ )
θ : Crank angle R at the closing timing of the intake valve 30 when the valve lift is 1 mm : Continuous ratio (connector rod length / crank radius)
It is.

By using the expression (2), the relationship between the effective compression ratio ε _r and the valve opening angle is converted into data based on the valve opening angle of the intake valve 30 when the valve lift is 1 mm, and used as a control map. Thus, the effective compression ratio ε _r can be precisely controlled.

  FIG. 23 is a graph showing an example of the ignition timing that is the basis of the control map in the embodiment of FIG.

Referring to FIG. 23, for example, if the geometric compression ratio of 11, ignition timing during normal operation, significant amounts A _Ig than the compression top dead center as shown by IG _a (e.g., the engine rotational speed 1500 rpm, crank angle CA = 6 ° to 8 °). In contrast, if the geometric compression ratio is 13, if the same knock characteristics as compression ratio 11 as shown by IG _v, but is was not ignited immediately before the compression top dead center, in this embodiment , as shown by IG _b, is set to be spark ignited at ignition timing is further retarded than the compression top dead center. Thereby, in this embodiment, in the retard operation region (low speed medium and high load operation region including the throttle full open region A _WOT ) A, the state in which the torque is not reduced and the high compression ratio (ε _r ≦ 14) is still maintained. (See FIG. 23).

When the ignition timing is retarded after compression top dead center, the retard amount R _Ig is experimentally accumulated in consideration of factors that determine knocking, such as in-cylinder temperature and in-cylinder pressure, and is determined by a control map. In this embodiment, for example, the retard amount R _Ig from the compression top dead center is set to a stroke range of 10% or less after the top dead center has elapsed (crank angle CA = around 35 ° after compression top dead center). By retarding the ignition timing after the compression top dead center, knocking is suppressed and operation at a high compression ratio is possible, but combustion is performed as much as the ignition timing is retarded from the compression top dead center. It is disadvantageous in terms of time. Therefore, in the present embodiment, in order to combust the fuel that has been able to suppress knocking and has shifted to the expansion stroke at an early stage, the retard amount R _{Ig is set} to a stroke range in which the piston 26 is 10% or less after the top dead center has elapsed. It is.

  FIG. 24 is a graph illustrating the fuel injection timing that is the basis of the control map in the embodiment of FIG. 14, where (A) is an example of split injection for forming a weakly stratified mixture, and (B) is stratified. It is an example of the division | segmentation injection for aiming at air-fuel | gaseous mixture formation.

  Referring to FIGS. 24 (A) and 24 (B), when the fuel injection timings F1 and F2 are divided into two or more times, fuel vaporization and atomization can be further promoted. It is possible to add turbulent flow and contribute to rapid combustion.

  Therefore, in this embodiment, fuel injection timings F1 and F2 are converted into data and used as a control map in accordance with various operating conditions.

  Here, in the example shown in FIG. 24A, the first half fuel injection timing F1 is executed, for example, in the middle of the intake stroke (for example, the end of fuel blow is 275 ° before compression top dead center). The timing F2 is set to the first half of the compression stroke (for example, the end of fuel blowing is 150 ° before the compression top dead center), and the fuel injection ratio of each timing F1 and F2 is set to 4/5: 1/5. . In this embodiment, the vaporization and atomization of the fuel injected in the intake stroke is promoted, and a weakly stratified mixture can be formed in the combustion chamber 27. Therefore, the combustion time can be shortened, and the high output and fuel consumption can be achieved. Can be improved.

  On the other hand, in the example shown in FIG. 24B, the first fuel injection timing F1 is executed in the first half of the compression stroke, for example, and the second fuel injection timing F2 is set in the middle or second half of the compression stroke. The fuel injection ratio of F2 is set to 4/5: 1/5. In this embodiment, since the latter half of the fuel injection can generate a large turbulent flow in the cylinder, even when the ignition timing is retarded, the combustion period can be shortened, and high output and fuel consumption can be obtained. It becomes possible. In this embodiment, since the concave portion 264 is formed in the central portion of the crown surface of the piston 26, it becomes possible to trap the fuel injected at the latter fuel injection timing F2, and ignite the stratified mixture. It can be formed in the retard operation region A. In the case of FIG. 24B, it is also possible to form a weak stratification at the ignition timing by adjusting the fuel injection amount at the fuel injection timings F1 and F2.

  25 and 26 are flowcharts showing a control flow according to the embodiment of FIG.

  First, referring to FIG. 25, in the above configuration, based on the input from the input elements (SW1 to SW8) to the engine main body 20 (step S10), the control unit 100 as the control means is in operation of the engine main body 20. It is determined whether or not there is (step S11).

  When the engine body 20 is stopped, the control unit 100 further determines whether or not there is a predetermined start request that is set in advance (for example, when the accelerator is depressed) (step S12), and the start is started. If no request is detected, the process returns to step S10 and waits. If there is a start request, the process proceeds to the next step S13.

  When the engine body 20 is in operation or when there is a start request, the control unit 100 further determines whether or not there is a rapid acceleration request from the low load operation region B (step S13). When there is a sudden acceleration request from the low-load operation region B, it is preferable to set the effective compression ratio high. Therefore, by using the control map M1 determined in advance by experiments or the like, the ignition timing is immediately after the compression top dead center. The maximum retarding value (crank angle CA = 35 °) is retarded (step S14).

On the other hand, when there is no sudden acceleration request in step S13, the control unit 100 further determines whether or not the operation region of the engine body 20 is operated in the ignition retard operation region A (step S15). If it is determined that the operation region of the engine body 20 is the ignition retard operation region A, the control unit 100 retards the closing timing of the intake valve 30 after the intake bottom dead center, as shown in FIG. The effective compression ratio ε _r is maintained at the geometric compression ratio (step S16). Next, the control unit 100 uses the control map M2 based on FIG. 23 to set the retard amount R _Ig of the spark plug 34 to a predetermined crank angle after compression top dead center (in this embodiment, the piston 26 is 10% or less after top dead center elapses). The stroke range) is set to start combustion early while suppressing knocking so that a large torque can be obtained (step S17).

After step S14 or step S17, the control unit 100 determines whether or not the operation state is capable of external EGR (step S18). This determination is performed by a known method by detecting / estimating the in-cylinder temperature and the in-cylinder pressure. However, in this embodiment, particularly in the case of the ignition retard operation region A including the throttle fully open region A _WOT. Is set to introduce an external EGR.

  If the operation state is such that external EGR can be introduced, the control unit 100 sets the EGR amount, the air-fuel ratio, and the fuel injection timing based on the control map M2 (step S19), and then operates the external EGR system 160. (Step S20), the engine body 20 is operated under the set conditions (Step S21). As a result, burned gas is introduced into the cylinder and the engine body 20 is operated at a low combustion temperature, so that heat loss can be reduced as much as possible. And after performing step S21, it returns to step S10 and repeats the control mentioned above.

  If it is determined in step S18 that the external EGR is not possible, the control unit 100 stops the external EGR system 160 (step S22), and sets the air-fuel ratio based on the control map M4 (step S23). The process proceeds to step S21.

  Next, the control when the operation region of the engine body 20 is not the ignition retard operation region A in step S15 will be described with reference to FIG.

As shown in FIG. 26, when the operation region of the engine body 20 is not the ignition retard operation region A, in the present embodiment, the control unit 100 performs in-cylinder operation based on the detection values of the engine water temperature sensor SW4 and the intake air temperature sensor SW2. the temperature T is estimated, and determines whether the in-cylinder temperature T is less than a predetermined reference value T _ST (step S24). If the in-cylinder temperature T is less than the reference value _TST , the control unit 100 sets the closing timing of the intake valve 30 based on the control map M5 and increases the effective compression ratio ε _r (step S25). ). That is, during cold weather, the vaporization and atomization of the spray is inadequate and the ignition delay of the end gas becomes large. Therefore, if the effective compression ratio ε _r remains low, the heat generation rate also remains low. Therefore, in the present embodiment, the geometric compression ratio that is set to be high is effectively used, and during a predetermined cold operation, the effective compression ratio ε _r is increased to stabilize combustion, and the output and fuel consumption are increased. We are trying to improve.

Further, the control unit 100 determines whether or not the set effective compression ratio ε _r is 12 or more (step S26). If the effective compression ratio ε _r is 12 or more, the control unit 100 proceeds to step S14. Then, the same operation as in the ignition retard operation region A is performed. This makes it possible to reliably prevent knocking when maintaining a high compression ratio. If the in-cylinder temperature T is equal to or higher than the reference value in step S24, the control unit 100 sets the closing timing of the intake valve 30 based on the control map M6, and lowers the effective compression ratio ε _r (step S27). Thereafter, or when it is determined in step S26 that the effective compression ratio ε _r is set to less than 12, the control unit 100 sets the ignition timing based on the control map M7 (step S28). Thereafter, it is determined whether or not internal EGR is possible (step S29). If possible, the control unit 100 sets the opening / closing timing of the exhaust valve, the air-fuel ratio, and the fuel injection timing based on the control map M8. (Step S30), the process proceeds to Step S21. If the operation region is incapable of executing internal EGR, the process proceeds to step S18 to determine whether external EGR is possible. In steps S19, S23, and S30 described above, in the low speed and low load operation region B1, the control maps M3, M4, M8 is set. In this embodiment, since the high expansion ratio can be secured by reducing the effective compression ratio under the high compression ratio condition, the fuel consumption can be sufficiently improved even in the operation at the stoichiometric air-fuel ratio. Therefore, the exhaust passage 153 is compared with the NOx catalyst. Since a three-way catalyst that is inexpensive and has a high purification rate can be arranged, it is possible to improve exhaust performance and fuel consumption even in a low-speed and low-load operation region.

As described above, in this embodiment, normally, in order to prevent knocking, in order to prevent knocking, a significant ignition timing retarded operation region (from the low speed and high load operation region A1 to the throttle fully open region) has been considered. In the middle and high load operation region A), the engine body 20 is operated while maintaining a high torque and fuel consumption with an effective compression ratio ε _r of 12 or more. That is, as shown in FIG. 23, the ignition timing IG _b to be retarded to avoid knock is, when it is set after the compression top dead center, as shown in FIG. 4, the piston 26 on the compression dead After the point elapses, the cool flame reaction in the cylinder becomes remarkable, and the combustion process after the compression top dead center elapses is a multistage ignition. As a result, the heat generation rate can be maintained while reducing the time loss, and sufficient Torque can be obtained. Further, by maintaining such a heat generation rate, the retard amount R _Ig can be reduced as much as possible. On the other hand, in the region where the cold flame reaction occurs, as shown in FIG. 5, as the number of moles increases, the in-cylinder temperature does not increase as much as the pressure increase. In addition, as shown in FIG. 6, the cold flame reaction occurs on the center side of the combustion chamber 27, and since there is little generation of end gas, an increase in the in-cylinder temperature is also suppressed. Under such temperature conditions, formaldehyde is generated, and the formaldehyde promotes consumption of OH radicals that cause knocking, and autoignition is also suppressed from this point. By constructing such a knocking suppression mechanism in the high compression ratio in the ignition retard operation region A (the high load operation region including the throttle fully open region A _WOT at least in the low speed region), the output decrease due to the ignition timing retard can be reduced. It is possible to compensate for the improvement in thermal efficiency and obtain as much fuel efficiency as a diesel engine without sacrificing output. Further, since the effective compression ratio ε _r is determined by the closing timing adjustment control of the intake valve 30, it is not necessary to use a complicated mechanism for changing the geometric compression ratio.

Further, in this embodiment, when the operation region of the engine body 20 is the low speed and low load operation region B1, the compression is performed by reducing the effective compression ratio ε _r obtained at the intake valve closing timing defined by the valve lift 1 mm to less than 12. The ignition plug 34 is ignited at a timing advanced by a predetermined amount from the top dead center, and the retard amount R _Ig from the compression top dead center of the ignition timing in the ignition retard operation region A is the same as that in the low speed and low load operation region B1. It is set smaller than the advance amount from the compression top dead center of the ignition timing. For this reason, in the present embodiment, in the case of the low speed and low load operation region B1, the effective compression ratio ε _r is lowered to less than 12 to prevent knocking in advance and the ignition timing is compressed in the same manner as a general engine. By advancing from the top dead center, it is possible to achieve proper combustion at a relatively high compression ratio corresponding to the operating region. Further, since the effective compression ratio ε _r is changed at the closing timing of the intake valve 30, it is possible to reduce the pumping loss and improve the fuel consumption. That is, when the intake valve 30 is closed late (or quickly closed) with an engine having a normal compression ratio, combustion becomes unstable as the effective compression ratio ε _r becomes considerably low. For this reason, there are many restrictions on the range in which late closing (or early closing) is possible, or there is a restriction that EGR cannot be sufficiently introduced. However, in this embodiment, since the geometric compression ratio is set to be considerably high, even if the effective compression ratio ε _r is considerably reduced, the actual compression ratio is still high, so that the combustion stability becomes high. Therefore, it is possible to widen the range of late closing (or early closing) of the intake valve 30 and, if the valve timing is the same, it is possible to increase the EGR rate as compared with the low compression ratio. . On the other hand, the retard amount R _Ig when the ignition timing is retarded in the ignition retard operation region A in the low speed region is set to a relatively small value. As a result, in the high load operation region in the low speed region, after shifting to the expansion stroke, it is possible to maintain extremely high torque while avoiding knocking.

Further, the low speed region in the present embodiment is a low speed region when the engine rotation region is divided into three stages of low speed, medium speed, and high speed, and a predetermined period (retard amount R _Ig ) in the low speed circuit is: The piston 26 has a stroke range of 10% or less after the top dead center has elapsed. For this reason, in this embodiment, the operating region is divided into three stages in the engine rotational speed region, and the intake valve 30 is closed so that the effective compression ratio ε _r is maintained at 12 or more in the throttle fully open region in the low rotational speed region. By adjusting the timing and retarding the ignition timing within a stroke range of 10% or less after the top dead center has elapsed, it is possible to achieve proper combustion at a relatively high compression ratio according to the operating region. become.

Further, in the present embodiment, the ignition timing is switched before the compression top dead center in the engine rotation range of medium speed or higher. In the present embodiment, on the high speed side after the medium speed, the effective compression ratio ε _r is appropriately reduced to less than 12 according to the operating state.

  Further, in the present embodiment, there is provided combustion period shortening means for shortening the combustion period of the air-fuel mixture when the ignition timing is retarded after compression top dead center. For this reason, in this embodiment, the combustion period shortening means can increase the heat generation rate in the expansion stroke as much as possible, suppress time loss, and obtain a high torque.

  FIG. 27 is a PV diagram related to the present embodiment.

  As shown in FIG. 27, when the ignition timing is retarded after compression top dead center, a time loss occurs as indicated by the phantom line, but by providing a combustion period shortening means (in the example shown, split injection is performed). When applied), the pressure increase after ignition can be accelerated and time loss can be reduced.

  As specific examples of the combustion period shortening means, the turbulent flow generating means for generating turbulent flow in the cylinder (the swirl-generating on-off valve 133a in FIGS. 14 and 15 and the reverse squish in FIG. 17) and FIG. It is possible to preferably employ split fuel injection. For this reason, in this embodiment, it is possible to increase the heat generation rate in the expansion stroke as much as possible, suppress time loss, and obtain a high torque by a relatively simple mechanism or control. Further, the combustion period shortening means may be provided with a plurality of spark plugs 34 for each cylinder, and the combustion period shortening means may be a multipoint ignition means for operating the plurality of spark plugs 34. In this aspect, since the combustion speed can be accelerated by multipoint ignition, the heat generation rate in the expansion stroke can be increased as much as possible, time loss can be suppressed, and high torque can be obtained.

In the present embodiment, an external EGR system 160 capable of adjusting the external EGR amount by the control unit 100 is provided, and the control unit 100 performs an ignition retard operation in which the operation region of the engine body 20 includes at least the throttle fully open region in the low speed region. In the case of the area A, an external EGR is introduced. For this reason, in the present embodiment, the combustion temperature can be lowered by the external EGR, so that cooling loss can be reduced and thermal efficiency is increased while knocking is avoided. As a result, high torque and fuel consumption can be obtained. That is, when the compression ratio is high, the in-cylinder temperature rapidly rises during the compression stroke, so that knocking is likely to occur. Furthermore, since the heat generated suddenly is absorbed by the wall surface of the cylinder and the like, the heat loss increases. On the other hand, when the burned gas discharged from the exhaust valve 31 is introduced, even when the effective compression ratio ε _r is in a relatively high state, the combustion temperature is lowered, and as a result, knocking and heat loss are suppressed. As a result, high torque and fuel consumption can be maintained.

In the present embodiment, the external EGR is also introduced in the low speed and low load operation region B1. For this reason, in this embodiment, it becomes possible to reduce heat loss as much as possible and maintain high fuel efficiency in combination with the decrease in the effective compression ratio ε _r .

Further, in this embodiment, at least in the low speed and low load operation region B1, the closing timing of the intake valve 30 is shifted from the intake bottom dead center by a predetermined amount so as to reduce the effective compression ratio εr. For this reason, in this embodiment, the effective compression ratio ε _r is reduced in the operation region where the combustion state tends to be relatively unstable, and a high expansion ratio is ensured. As a result, it is possible to reduce the pumping loss and improve fuel efficiency while preventing knocking due to the high compression ratio.

Further, in the present embodiment, the external EGR system 160 serving as EGR means for introducing the E GR into the cylinder is provided, the external EGR system 160 is one in which at least said at low speed and low load operation region B1, introducing EGR. For this reason, in this embodiment, it becomes possible to reduce a heat loss as much as possible and to maintain a high fuel consumption in combination with a decrease in the effective compression ratio. That is, when the compression ratio is high, the in-cylinder temperature rapidly increases during the compression stroke. Here, since the heat generated suddenly is absorbed by the wall surface of the cylinder and the like, the heat loss increases. On the other hand, when the burned gas discharged from the exhaust valve is introduced, the effective compression ratio is lowered, and as a result, the combustion temperature is lowered, so that heat loss is suppressed and high fuel consumption is maintained. It becomes possible to do.

  In the present embodiment, the air-fuel ratio is set to the stoichiometric air-fuel ratio in the low speed and low load operation region B1. As described above, since the high expansion ratio can be secured by reducing the effective compression ratio under the high compression ratio condition, the fuel consumption can be sufficiently improved even in the operation at the theoretical air-fuel ratio. Therefore, the exhaust passage 153 is compared with the NOx catalyst. Since a three-way catalyst having a low purification rate and a high purification rate can be arranged, sufficient exhaust performance can be achieved even in the low speed and low load operation region.

  In the present embodiment, the low speed and low load operation region B1 includes the idling operation region B2. For this reason, in this embodiment, it becomes possible to maintain high fuel efficiency even in the idling operation region B2 where the usage frequency is high.

In the present embodiment, in-cylinder temperature estimation means for estimating the in-cylinder temperature of the engine main body 20 is provided, and the control unit 100 sets the closing timing of the intake valve 30 to the vicinity of the bottom dead center of intake during cold start. The closing timing of the intake valve 30 is adjusted and controlled so as to increase the effective compression ratio ε _r and ensure sufficient intake. For this reason, in the present embodiment, the volumetric efficiency can be increased by increasing the effective compression ratio ε _r and ensuring sufficient intake air, so that it is sufficient for raising good ignition / combustion performance and engine speed. It is possible to obtain a proper torque.

In the present embodiment, the accelerator opening sensor SW6 is provided as an engine acceleration detecting means for detecting engine acceleration, and the control unit 100 sets the ignition timing after compression top dead center at the time of sudden acceleration from the low load operation region. In this case, the maximum retarded value of the predetermined period (retard amount R _Ig ) is retarded at once. For this reason, in the present embodiment, knocking due to high temperature fresh air sucked during rapid acceleration is avoided.

In this embodiment, the fuel injection valve 32 is a direct injection type that injects fuel toward the vicinity of the electrode of the spark plug 34, and the crown surface peripheral portion of the piston 26 of the engine body 20 is provided on the crown surface. And a ridge 33 for generating reverse squish flow Ra, Rb when moving from the compression top dead center to the expansion stroke, and a recess 264 formed in the central portion of the crown surface. 100 controls the fuel injection valve 32 to inject fuel in the compression stroke. For this reason, in the present embodiment, in the process leading to the compression stroke, the flight space of the injected fuel is secured by the concave portion 264 formed in the central portion of the crown surface of the piston 26. The reverse squish flows Ra and Rb are formed in the peripheral portion of 26. As a result, the combustion period is shortened, knocking prevention and heat generation rate in the expansion stroke are increased as much as possible, time loss is suppressed, and high torque and fuel efficiency are improved. Further, in the operation region where the effective compression ratio ε _r is operated at 12 or more, the concave portion 264 contributes to the generation of the cool flame in the combustion chamber 27 after the compression top dead center has elapsed, and it becomes a factor that further increases the output.

As described above, the present embodiment conventionally employs an expensive mechanism or reduces the effective compression ratio ε _r at the closing timing of the intake valve 30 to cope with the output at the high load operation region A. Since knocking is avoided while maintaining a high compression ratio, there is a remarkable effect that fuel efficiency equivalent to that of a diesel engine can be obtained as much as possible while combining low cost and high output.

  The above-described embodiments are merely preferred specific examples of the present invention, and the present invention is not limited to the above-described embodiments.

For example, in the embodiment described above, the engine body 20 is preferably operated using a fuel having an octane number of 96 RON or more. In that case, in the ignition retard operation region A including the throttle fully open region A _WOT in the low speed region, the effective compression ratio ε _{r is set} to 12 or more, and the ignition timing IG _b is set to a predetermined retard amount R after the compression top dead center has elapsed. By retarding only _Ig , a high torque can be obtained most effectively utilizing the cold flame reaction in the cylinder. As described with reference to FIGS. 3 and 11, when fuel of 96 RON or more is injected, the inside of the cylinder becomes equal to or higher than the activation energy that causes the cold flame reaction when the compression ratio is 12 or more, and the cold flame reaction is caused by the ignition retard. It is possible to improve the amount of heat generated by the heat and increase the torque.

In the embodiment described above, the upper limit of the geometric compression ratio ε of the engine body 20 is preferably 15. In that case, pre-ignition can be achieved even if a high effective compression ratio ε _r is maintained under conditions where auto-ignition is likely to occur, such as during low-speed full-load operation where the intake air temperature is high or when the engine is warm when restarting. Etc. can be prevented.

Further, as another aspect of the present invention, when the engine body 20 is operated with fuel of 91 RON or more, the geometric compression ratio is set to 12.5 or more, and the ignition including the throttle fully open region A _WOT in the low speed region. During the operation in the retard operation region A, the intake valve closing timing is adjusted so as to maintain the effective compression ratio ε _r obtained at the intake valve closing timing defined by the valve lift of 1 mm at 11.5 or more, and the ignition timing IG _b is set. You may comprise so that it may retard within the predetermined period after a compression top dead center. In such an aspect, even when a fuel having a relatively low octane number is used, the cold flame reaction in the cylinder is effectively utilized in the ignition retard operation region A including the throttle fully open region A _WOT in the low speed region, High torque can be obtained.

  In a spark ignition gasoline engine that is operated using a fuel having an octane number of 91 RON or more, the upper limit of the geometric compression ratio of the engine body 20 is preferably 14.5. In such a case, pre-ignition and the like are prevented from occurring even if a high effective compression ratio ε is maintained under conditions where auto-ignition is likely to occur, such as when the intake air temperature is high or when the engine is warm. be able to.

  Furthermore, as another aspect of the present invention, when the engine body 20 is operated using fuel of 100 RON or more, the upper limit of the geometric compression ratio of the engine body 20 is 15.5. Is preferred. In such a case, pre-ignition and the like are prevented from occurring even if a high effective compression ratio ε is maintained under conditions where auto-ignition is likely to occur, such as when the intake air temperature is high or when the engine is warm. be able to.

Further, as a method of reducing the effective compression ratio ε _r , in the above-described embodiment, the valve operating mechanism 40 that can change the opening / closing timing of the intake valve 30 in a stepless manner is used. A valve mechanism with a so-called lost motion function that can switch the opening / closing timing of the intake valve 30 in two stages by selectively transmitting the cam to the intake valve 30 may be used.

  FIG. 28 is a graph showing a control example using a valve mechanism with a lost motion function.

As shown in FIG. 28, when a valve mechanism with a lost motion function is used, the closing timing of the intake valve 30 is retarded. In this embodiment, since the air once introduced into the cylinder is pushed out, the pumping loss slightly occurs, but it is possible to reduce the effective compression ratio ε _r with an inexpensive mechanism and avoid knocking.

  Further, the valve mechanism with the lost motion function as described above may be adopted as means for executing the internal EGR.

  FIG. 29 is a configuration diagram showing a configuration of an intake air heating system as intake air heating means according to another embodiment of the present invention.

  Referring to FIG. 29, when so-called premixed compression auto-ignition combustion (HCCI) is performed by employing the present invention, a heater as intake air heating means as shown in FIG. 140 is preferably provided.

  More specifically, a three-way electromagnetic valve 138 is provided upstream of the throttle valve 137 in the intake passage 136, and a heater 140 is provided in a bypass passage 139 connected to the three-way electromagnetic valve 138. Further, the heater 140 is provided with a temperature sensor SW7 so that the temperature of the intake air in the bypass passage 139 heated by the heater 140 can be detected. The temperature sensor SW7 is connected to a control unit (not shown).

  FIG. 30 is a configuration diagram showing the configuration of the intake air heating system 170 as the intake air heating means according to the embodiment of FIG.

  Referring to FIG. 30, a heating passage 171 is branched and connected to the intake passage 136. A cooling water heat exchanger 172 and an exhaust heat exchanger 173 are connected in the middle of the heating passage 171.

  The heating passage 171 is for returning the heat absorbed through the heat exchangers 172 and 173 to the intake side. A branch pipe 174 a branched for each cylinder 24 is provided on the downstream side of the heating passage 171, and each branch pipe 174 a is connected to a port on the intake side of the corresponding on-off valve 134.

  The cooling water heat exchanger 172 is connected to the water cooling system 174 of the engine body 20 and absorbs the heat absorbed by the cooling water returning from the engine body 20 to the radiator (not shown) into the intake air passing through the heating passage 171. Is for.

  The exhaust heat exchanger 173 is connected to the exhaust passage 153 of the engine body 20 and absorbs the heat of burned gas into the intake air passing through the heating passage 171. The exhaust heat exchanger 173 is disposed on the downstream side of the cooling water heat exchanger 172 in the heating passage 171.

  In the present embodiment, these heat exchangers 172 and 173 constitute the main part of the intake air heating system 170.

  In this configuration, the control unit 100 controls the three-way solenoid valve 138 so that the valve opening ratio can be changed in the same manner as the open / close valve 134. By switching the three-way solenoid valve 138, the outside air is freshened. The air can be introduced into the intake manifold 132 as it is, or the air heated by the heater 140 can be introduced into the intake manifold 132.

  It goes without saying that various modifications can be made within the scope of the claims of the present invention.

It is a graph which shows an example of the ignition retard at the time of high load driving | operation. It is a graph which shows the relationship between the crank angle and torque for demonstrating the hypothesis in the development process of this invention. It is a graph which shows the simulation result which shows the relationship between ignition timing and IMEP. 6 is a graph showing a relationship between a heat generation rate and a crank angle when an engine having a compression ratio of 11, 13, 14, 15 is ignited at 8 ° CA after the compression top dead center has elapsed. It is the graph which simulated the combustion process after compression top dead center progress by a high compression ratio, the upper stage shows the relationship between pressure and time, and the lower stage shows the relationship between the number-of-moles increase ratio and time. It is a contour line which shows the temperature distribution of a combustion chamber when the compression top dead center is reached. It is a graph which shows the in-cylinder pressure at the time of combustion and the adiabatic compression temperature history of the end gas part of a peripheral part, the upper stage shows the relationship between a pressure and a crank angle, and the lower stage shows the relationship between an end gas temperature and a crank angle. It is a graph which shows the relationship between the heat release rate in case a compression ratio is 14, and a crank angle. It is a PV diagram at the time of the compression ratio 14 based on numerical simulation. It is a graph which shows the relationship between a compression ratio and the emitted-heat amount by a cold flame reaction for every octane number. It is a graph which shows the relationship between the compression ratio calculated based on the graph of FIG. 10, and the indicated mean effective pressure (IMEP) for every octane number. It is a graph which shows the relationship between the geometric compression ratio at the time of employ | adopting a variable valve timing system, and an effective compression ratio. It is a graph which shows the distribution of ignition delay by the relationship between pressure and temperature. It is a block diagram which shows schematic structure of the control apparatus which concerns on one Embodiment of this invention. FIG. 15 is a schematic sectional view showing a structure of one cylinder of the four-cycle spark ignition type gasoline engine according to FIG. 14. It is a plane schematic diagram expanding and showing a cylinder. It is explanatory drawing which shows the airflow of the combustion chamber which concerns on this embodiment, (A) has shown the compression stroke initial stage, (B) has shown the expansion stroke initial stage, respectively. It is a perspective view which shows the specific structure of the valve mechanism based on embodiment of FIG. FIG. 19 is a cross-sectional view showing the main part of the valve mechanism of FIG. 18, where (A) shows when the lift amount is 0 in the large lift control state, and (B) shows when the lift amount is maximum in the large lift control state. (C) shows when the lift amount is 0 in the small lift control state, and (D) shows when the lift amount is maximum in the small lift control state. FIGS. 19B and 19D schematically represent the control states, where FIG. 19A corresponds to the large lift control position, and FIG. 19B corresponds to the small lift control position. It is a graph which shows the relationship between the engine speed used as the basis of a control map in the embodiment of FIG. 14, and a request torque. It is a timing chart which shows the example of control of the effective compression ratio used as the basis of a control map in embodiment of FIG. It is a graph which shows an example of the ignition timing used as the basis of a control map in embodiment of FIG. FIG. 15 is a graph illustrating fuel injection timing that is a basis of a control map in the embodiment of FIG. 14, where (A) is an example of split injection for forming a weakly stratified mixture, and (B) is a stratified mixture formation. It is an example of the division | segmentation injection for aiming at. It is a flowchart which shows the control flow which concerns on embodiment of FIG. It is a flowchart which shows the control flow which concerns on embodiment of FIG. It is a PV diagram relevant to this embodiment. It is a graph which shows the example of control using the valve mechanism with a lost motion function. It is a block diagram which shows the structure of the intake-air heating system as an intake-air heating means which concerns on another embodiment of this invention. It is a block diagram which shows the structure of the intake-air heating system as an intake-air heating means which concerns on embodiment of FIG.

10 Four-cycle spark ignition gasoline engine 20 Engine body 21 Crankshaft 22 Cylinder block 23 Cylinder head 24 Cylinder 26 Piston 27 Combustion chamber 28 Intake port 29 Exhaust port 30 Intake valve 31 Exhaust valve 32 Fuel injection valve 33 Raised portion 34 Spark plug 35 Ignition circuit 40 Valve mechanism 100 Control unit (an example of control means)
160 External EGR system 170 Intake heating system 264 Recess A Ignition retard operation region B Normal ignition operation region B1 Low speed low load operation region B2 Idling operation region F1, F2 Fuel injection timing N Engine rotation speed Ra, Rb Reverse squish flow R _Ig retard amount S Offset SW1-SW7 Sensor T In-cylinder temperature Ta1 Normal tumble flow Ta2 Reverse tumble flow T _ST reference value ε _r Effective compression ratio

Claims (22)

In a four-cycle spark ignition gasoline engine with at least a spark plug,
An engine body with a geometric compression ratio set to 13 or higher;
An intake valve and an exhaust valve which are respectively provided in an intake port and an exhaust port connected to a cylinder of the engine body, and open and close corresponding ports;
A supercharger provided in an intake passage for introducing fresh air into the intake port;
Driving state detecting means for detecting the driving state of the engine body;
Control means for performing at least the ignition timing adjustment control of the ignition plug, the effective compression ratio adjustment control by the intake valve closing timing adjustment control, and the supercharger drive control based on the detection of the operating state detection means And when the operating region of the engine body is a high load operating region including at least the throttle fully open region in the low speed region, the control means is obtained at the intake valve closing timing defined by the valve lift of 1 mm. By adjusting the intake valve closing timing to maintain the effective compression ratio at 12 or more and retarding the ignition timing within a predetermined period after the compression top dead center, a cold flame reaction is caused after the compression top dead center elapses. The spark-ignition gasoline engine is characterized in that the air-fuel mixture is burned by flame propagation based on spark ignition. The spark ignition gasoline engine according to claim 1,
The spark-ignition gasoline engine is characterized in that the engine body is operated using a fuel having an octane number of 96 RON or more. The spark ignition gasoline engine according to claim 2,
The upper limit of the geometric compression ratio of the engine body is 15. A spark ignition gasoline engine. The spark ignition gasoline engine according to claim 1,
The engine body is operated using a fuel having an octane number of 100 RON or more,
The upper limit of the geometric compression ratio of the engine body is 15.5. A spark ignition gasoline engine. In a four-cycle spark ignition gasoline engine with at least a spark plug,
An engine body operated with fuel having a geometric compression ratio of 12.5 or higher and an octane number of 91 RON or higher;
An intake valve and an exhaust valve which are respectively provided in an intake port and an exhaust port connected to a cylinder of the engine body, and open and close corresponding ports;
A supercharger provided in an intake passage for introducing fresh air into the intake port;
Driving state detecting means for detecting the driving state of the engine body;
Control means for performing at least the ignition timing adjustment control of the ignition plug, the effective compression ratio adjustment control by the intake valve closing timing adjustment control, and the supercharger drive control based on the detection of the operating state detection means And when the operating region of the engine body is a high load operating region including at least the throttle fully open region in the low speed region, the control means is obtained at the intake valve closing timing defined by the valve lift of 1 mm. By adjusting the intake valve closing timing so that the effective compression ratio is maintained at 11.5 or more and retarding the ignition timing within a predetermined period after the compression top dead center, a cold flame reaction occurs after the compression top dead center elapses. A spark-ignited gasoline engine characterized in that the air-fuel mixture is burned by flame propagation based on spark ignition. The spark ignition gasoline engine according to claim 5,
The upper limit of the geometric compression ratio of the engine body is 14.5. A spark ignition gasoline engine. The spark ignition type gasoline engine according to any one of claims 1 to 4,
Wherein when the operating region of the engine body is of low speed and low load operation region, from the compression top dead center is lowered the effective compression ratio to less than 12 Ru ignites the ignition plug at a predetermined amount advanced the timing This is a spark ignition gasoline engine. The spark ignition gasoline engine according to claim 7 ,
The spark ignition gasoline engine according to claim 1, wherein the predetermined period is set to be smaller than an advance amount from a compression top dead center of an ignition timing in the low speed and low load operation region . The spark ignition gasoline engine according to claim 1, 2, 3, 4, 7 or 8 .
The low speed range set in the control means is a low speed range when the engine rotation range is divided into three stages of low speed, medium speed, and high speed, and the predetermined period is after the top dead center has elapsed. A spark ignition gasoline engine characterized by a stroke range of 10% or less . The spark ignition gasoline engine according to claim 9 ,
The spark igniting gasoline engine characterized in that the control means switches the ignition timing before the compression top dead center in an engine speed range of medium speed or higher . The spark ignition gasoline engine according to any one of claims 1 to 10 ,
A spark ignition type gasoline engine comprising combustion period shortening means for shortening a combustion period of an air-fuel mixture when ignition timing is retarded after compression top dead center . The spark ignition gasoline engine according to claim 11 ,
The spark ignition type gasoline engine, wherein the combustion period shortening means is turbulent flow generating means for generating turbulent flow in a cylinder . The spark ignition gasoline engine according to claim 11 ,
A plurality of spark plugs are provided for each cylinder,
The spark ignition type gasoline engine, wherein the combustion period shortening means is multi-point ignition means for operating a plurality of spark plugs . The spark ignition gasoline engine according to any one of claims 1 to 13,
A fuel injection valve capable of controlling the injection timing by the control means;
When the operating range of the engine body is at least a predetermined medium load operating range from a predetermined middle load operating range to a throttle full open range in the low speed range, fuel is supplied within a predetermined period of time from the intake stroke to the compression stroke. A spark-ignited gasoline engine characterized by executing split injection for multiple injections . The spark ignition gasoline engine according to any one of claims 1 to 14 ,
An external EGR system capable of adjusting the external EGR amount by the control means is provided,
The spark-ignited gasoline engine is characterized in that the control means introduces an external EGR when the operating range of the engine body is a high-load operating range including a throttle full open range at least in a low speed range . The spark ignition gasoline engine according to claim 15 ,
The spark ignited gasoline engine characterized in that the control means introduces an external EGR at least in a low speed and low load operation region . The spark ignition type gasoline engine according to any one of claims 1 to 16 ,
The spark-ignited gasoline is characterized in that the control means shifts the closing timing of the intake valve from the intake bottom dead center by a predetermined amount so as to reduce the effective compression ratio at least in a low-speed and low-load operation region. engine. The spark ignition gasoline engine according to claim 17,
EGR means for introducing EGR into the cylinder is provided,
The spark-ignited gasoline engine , wherein the EGR means introduces EGR at least in the low-speed and low-load operation region . The spark ignition gasoline engine according to claim 18 ,
The spark-ignition gasoline engine according to claim 1, wherein the control means sets the air-fuel ratio to the stoichiometric air-fuel ratio in the low-speed and low-load operation region . The spark ignition gasoline engine according to any one of claims 17 to 19 ,
The spark-ignition gasoline engine, wherein the low-speed and low-load operation region set in the control means includes an idling operation region . The spark ignition type gasoline engine according to any one of claims 1 to 20 ,
In-cylinder temperature estimating means for estimating the in-cylinder temperature of the engine body,
The control means sets the intake valve closing timing in the vicinity of the intake bottom dead center during cold start, and adjusts and controls the intake valve closing timing so as to increase the effective compression ratio and ensure sufficient intake. spark-ignition gasoline engine, characterized in that. The spark ignition gasoline engine according to any one of claims 1 to 21,
A fuel injection valve capable of controlling the injection timing by the control means;
The fuel injection valve is a direct injection type that injects fuel toward the vicinity of the electrode of the spark plug,
A piston crown surface of the engine body is formed in a peripheral portion of the crown surface, and is formed in a ridge portion that generates a reverse squish flow when transitioning from a compression top dead center to an expansion stroke, and a central portion of the crown surface. A recess is provided,
The spark igniting gasoline engine characterized in that the control means controls a fuel injection valve so as to inject fuel in a compression stroke.