Internal combustion engine, control or regulating device therefor, method for controlling an engine and computer program product

Abstract

Internal combustion engine, including: a combustion chamber (17) configured to cause combustion of a gas mixture within the combustion chamber (17); a spark plug (25) arranged to be oriented substantially into the combustion chamber (17) and configured to ignite the gas mixture within the combustion chamber (17); and a controller (10) which is operationally connected to the spark plug (25), wherein the combustion is carried out in at least one given mode (SPCCI (Spark Controlled Compression Ignition) combustion) in which spark ignition combustion (SI combustion) and compression ignition combustion (CI combustion) are combined, in which, after the spark plug (25) has ignited the gas mixture to start the combustion, unburned gas mixture is combusted by auto-ignition, wherein the controller (10) has a heat quantity ratio change module (102a) configured to change a heat quantity ratio as an index relating to a ratio of the heat quantity generated by SI combustion to a total heat quantity generated by SPCCI combustion, according to an operating state of the engine, and wherein the controller (10) causes the heat quantity ratio change module (102a) to increase the heat quantity ratio at a high motor speed compared to a low motor speed.

Description

TECHNICAL FIELD

[0001] The present disclosure relates to an internal combustion engine operable in at least one given mode in which, after a spark plug ignites the gas mixture to initiate combustion, unburned gas mixture combusts by auto-ignition. Furthermore, the invention relates to a control system or device for a compression-ignition engine, a method for controlling an engine, and a computer program product. BACKGROUND OF REVELATION

[0002] JP 4 082 292 B2 discloses an engine for combusting a gas mixture within a combustion chamber by auto-ignition within a given operating range in which engine load and engine speed are low. The engine combusts the gas mixture by spark ignition within an operating range in which the engine load is higher and an operating range in which the engine speed is higher than the given operating range.

[0003] Furthermore, the combustion caused by compression ignition is accompanied by relatively loud combustion noise. At high engine speeds, the engine's NVH (noise vibration harshness) exceeds an acceptable level.

[0004] DE 10 2010 009 817 A1 describes an operating strategy for HCCI combustion during engine warm-up. The internal combustion engine comprises a combustion chamber configured to effect combustion of a gas mixture within the combustion chamber, a spark plug arranged to be substantially aligned into the combustion chamber and configured to ignite the gas mixture within the combustion chamber, and a controller operatively connected to the spark plug. Combustion is performed in at least one given mode in which, after the spark plug has ignited the gas mixture to initiate combustion, unburned gas mixture combusts by auto-ignition.

[0005] DE 10 2011 009 247 A1 discloses a method for operating an engine with spark-assisted compression ignition and direct injection. SUMMARY OF REVELATION

[0006] The present disclosure has been made in view of the above situations and aims to perform compression ignition combustion while suppressing NVH of a compression ignition engine below an allowable value.

[0007] This problem is solved by the features of the independent claims. Further developments are defined in the dependent claims.

[0008] As described later, the inventors of the present invention considered a combustion mode (SPCCI (Spark Controlled Compression Ignition) combustion) in which SI (Spark Ignition) combustion and CI (Compression Ignition) combustion are combined. That is, a gas mixture within a combustion chamber is forcibly ignited to burn by flame propagation, and heat generated by this combustion causes unburned gas mixture to burn by self-ignition.

[0009] In CI combustion, the timing of auto-ignition changes significantly due to a change in the temperature within the combustion chamber before compression begins. For example, if the auto-ignition timing is advanced, combustion noise increases.

[0010] In this regard, the change in the temperature within the combustion chamber before the start of compression can be reduced by changing the amount of heat generated during SI combustion. For example, by changing the ignition timing to change the start time of SI combustion according to the temperature within the combustion chamber before the start of compression, the timing of auto-ignition can be controlled.

[0011] Flame propagation combustion causes relatively small pressure fluctuations, thus reducing combustion noise. Furthermore, CI combustion has a shorter combustion duration compared to flame propagation combustion, which is beneficial for improving fuel efficiency. Therefore, the combustion mode that combines SI combustion and CI combustion improves fuel efficiency while reducing combustion noise.

[0012] By performing SPCCI combustion when an engine speed is high, it is possible to perform CI combustion while suppressing NVH below the allowable value.

[0013] In particular, according to one aspect of the present disclosure, an internal combustion engine and a control system therefor are provided. The engine includes a combustion chamber configured to cause combustion of a gas mixture within the combustion chamber, a spark plug arranged to be substantially aligned into the combustion chamber and configured to ignite the gas mixture within the combustion chamber, and a controller operatively connected to the spark plug and configured to operate the engine by outputting a control signal to the spark plug.

[0014] Combustion is performed in at least one given mode (SPCCI (Spark Controlled Compression Ignition) combustion), which combines spark ignition (SI) combustion and compression ignition (CI) combustion, in which, after the spark plug ignites the gas mixture to begin combustion, unburned gas mixture combusts through self-ignition. The controller includes a heat ratio changing module configured to change, according to an operating state of the engine, a heat ratio as an index relating to a ratio of the heat amount generated by the SI combustion with respect to a total heat amount generated by the SPCCI combustion. The controller causes the heat ratio changing module to increase the heat ratio at a high engine speed compared to a low engine speed.

[0015] It should be noted that the definition of “combustion chamber” here is not limited to a space formed when a piston is at top dead center on a compression stroke (CTDC), but is wide.

[0016] According to this configuration, the controller, when determined as needed based on an engine operating condition, outputs the control signal to the spark plug to perform combustion in the given mode, in which the spark plug ignites the gas mixture within the combustion chamber at a given time to initiate combustion, and then the unburned gas mixture combusts through self-ignition. That is, the SPCCI combustion described above is performed.

[0017] To control SPCCI combustion, the heat ratio (SI ratio) is defined in the controller as an index that indicates the characteristics of this combustion. Furthermore, the controller includes the heat ratio modification module, which is configured to modify the heat ratio according to the engine's operating state.

[0018] When SPCCI combustion is performed, the heat ratio changing module controls SPCCI combustion such that the heat ratio becomes higher at a high engine speed than at a low engine speed.

[0019] As described above, NVH increases at higher engine speeds than at lower engine speeds. When CI combustion, which is accompanied by loud combustion noise, is performed at high engine speeds, NVH may exceed the allowable level, similar to when the engine speed is low.

[0020] Therefore, when the engine speed is high, by controlling the heat quantity ratio to be higher, that is, relatively increasing the SI combustion ratio, which is accompanied by a quieter combustion noise in SPCCI combustion, to relatively decreasing the CI combustion ratio, NVH is suppressed below the allowable value even at the high engine speed.

[0021] As a more specific configuration, the engine and / or the control system may further include a pressure sensor configured to detect pressure in the combustion chamber. The controller may set a target heat ratio, receive a detection signal from the pressure sensor, calculate the heat ratio based on a pressure waveform caused by the combustion of the gas mixture, and, if the calculated heat ratio differs from the target heat ratio, change the heat ratio to approximate the target heat ratio.

[0022] Thus, the heat quantity ratio is changed according to the difference between the actual combustion state within the combustion chamber based on the pressure sensor's detection signal and the target combustion state. Since the target combustion state of the SPCCI combustion is accurately achieved according to the engine's operating state, NVH is reduced below the permissible level.

[0023] The engine or control system may further include an internal exhaust gas recirculation (EGR) system configured to change an internal EGR ratio, which is a ratio of an amount of internal EGR gas contained in the gas mixture within the combustion chamber. The controller may output a control signal to the internal EGR system to increase the internal EGR ratio to increase the heat quantity ratio.

[0024] One method for increasing the heat quantity ratio at a high engine speed is to increase a temperature of the gas mixture within the combustion chamber (a temperature immediately before combustion).

[0025] At high engine speed, a combustion cycle is shorter than at low engine speed. Since the amount of heat absorbed by an inner wall of the combustion chamber, etc., during combustion decreases accordingly, and the temperature within the combustion chamber decreases, it becomes difficult for the gas mixture to absorb heat, and a period for combustion of the gas mixture becomes short. Therefore, at high engine speed, the heat quantity ratio must be increased under such adverse conditions. In this regard, raising the temperature of the gas mixture itself stimulates SPCCI combustion even under such conditions.

[0026] In SPCCI combustion, ignition of CI combustion occurs after SI combustion. Although increasing the gas mixture temperature stimulates both SI combustion and CI combustion, the combustion duration is short when the engine speed is high. Thus, SI combustion becomes sharper and is stimulated more quickly than CI combustion, which starts later than SI combustion. Thus, increasing the gas mixture temperature increases the ratio of SI combustion at high engine speeds.

[0027] Therefore, by increasing the internal EGR ratio to increase the ratio of internal EGR gas contained within the gas mixture (high-temperature combustion gas), the temperature of the gas mixture increases, thus increasing the heat ratio. As a result, NVH is reduced below the permissible level even at high engine speeds.

[0028] Furthermore, the engine or control system may further include an external EGR system configured to change an external EGR ratio, which is a ratio of the amount of external EGR gas contained in the gas mixture within the combustion chamber. The controller may output a control signal to the external EGR system to decrease the external EGR ratio to thereby increase the heat quantity ratio.

[0029] While the engine is running, reducing the external EGR ratio to reduce the ratio of external EGR gas contained in the gas mixture (cooled low-temperature combustion gas) also increases the temperature of the gas mixture and thus increases the heat ratio. As a result, NVH is reduced below the allowable level even at high engine speeds.

[0030] By increasing the internal EGR ratio and decreasing the external EGR ratio, the temperature of the gas mixture before combustion is increased more effectively, and thus the heat quantity ratio is increased in a stable manner.

[0031] A swirl flow can be formed in the combustion mode when the combustion is carried out in the given mode.

[0032] When the swirl flow is formed in the combustion chamber, the flow of the gas mixture within the combustion chamber is stimulated and the gas mixture in the combustion chamber is also stratified. This prevents dangerous substances such as NO xand soot are generated, and the gas mixture that stably performs SI combustion in SPCCI combustion is formed around the spark plug at the ignition timing. Therefore, the degree of freedom of controlling the heat quantity ratio is increased, and more stable SPCCI combustion is achieved.

[0033] The controller can output the control signal to the spark plug to advance the ignition timing to increase the heat quantity ratio.

[0034] When the ignition timing is advanced, SI combustion begins early. Thus, SI combustion is stimulated even in a short period of time at high engine speeds, and the heat ratio is increased. This method can be combined with the methods described above, and the heat ratio is thereby increased more effectively at high engine speeds.

[0035] The internal EGR gas can be introduced into the combustion chamber by overlapping an open period of an intake valve with an open period of an exhaust valve.

[0036] When introducing internal EGR gas by setting a so-called positive overlap period, in which the intake valve opening period and the exhaust valve opening period overlap, the high-temperature combusted gas generated in the combustion chamber flows out of the combustion chamber into an intake tract. Then, the combusted gas that has flowed into the intake tract is re-introduced into the combustion chamber. By flowing to the intake tract where the temperature is lower, the high-temperature combusted gas to be introduced into the combustion chamber—that is, the internal EGR gas—is cooled.

[0037] Therefore, introducing internal EGR gas by setting a positive overlap period lowers the combustion chamber temperature more than introducing internal EGR gas by setting a negative overlap period, during which the high-temperature combusted gas is confined as it is in the combustion chamber. This temperature reduction allows a larger amount of internal EGR gas to be introduced, thus slowing down CI combustion by auto-ignition. As a result, the degree of freedom of controlling the heat quantity ratio is increased, and stable SPCCI combustion is achieved.

[0038] A swirl ratio of the swirl flow can be less than 4.

[0039] In this case, the control system may, for example, further include a swirl control valve provided on the engine and configured to change the intensity of the swirl flow by controlling an opening through the controller. The intensity of the swirl flow may increase as the opening of the swirl control valve is reduced. The controller may control the swirl control valve to have a given narrow opening without completely closing the swirl control valve.

[0040] Although the formation of swirl flow in the combustion chamber stimulates the flow of the gas mixture within the combustion chamber and stratifies the gas mixture as described above, if the swirl flow is excessively strong, the stratification of the gas mixture may be unstable depending on the combustion conditions. When the swirl ratio is below approximately 4, the flow of the gas mixture within the combustion chamber becomes suitable, and thus, stable stratification of the gas mixture is achieved over a wide operating range of the engine.

[0041] In particular, according to another aspect of the present disclosure, an internal combustion engine and a control system therefor are provided. The engine includes a combustion chamber configured to cause combustion of a gas mixture within the combustion chamber, a spark plug arranged to be substantially aligned into the combustion chamber and configured to ignite the gas mixture within the combustion chamber, an EGR system configured to change an EGR ratio, which is a ratio of an amount of EGR gas contained in the gas mixture within the combustion chamber, and a controller connected to the spark plug and the EGR system and configured to operate the engine by outputting a control signal to the spark plug and the EGR system, respectively.

[0042] The controller outputs the control signal to the spark plug at a given ignition timing to perform combustion in a given mode, in which, after the spark plug ignites the gas mixture to initiate combustion, the unburned gas mixture in the combustion chamber combusts through auto-ignition. The controller outputs the control signal to the EGR system to increase the EGR ratio more at high engine speeds than at low engine speeds.

[0043] According to this configuration, the controller, when determined as necessary based on the engine's operating condition, outputs the control signal to the spark plug to perform combustion in the given mode, in which the spark plug ignites the gas mixture within the combustion chamber at a given time to initiate combustion, and then the unburned gas mixture combusts through self-ignition. That is, the SPCCI combustion described above is performed.

[0044] When SPCCI combustion is performed, the controller outputs the control signal to the EGR system provided on the engine. The EGR system changes the EGR ratio (a ratio of the amount of EGR gas contained in the gas mixture within the combustion chamber) and increases the EGR ratio to be higher at high engine speed than at low engine speed, according to the control signal from the controller.

[0045] As described above, NVH increases at high engine speeds compared to low engine speeds. When CI combustion, which is accompanied by loud combustion noise, is performed at high engine speeds, NVH may exceed the allowable level, similar to when the engine speed is low. Therefore, to perform auto-ignition combustion at high engine speeds, it is necessary to reduce the CI combustion ratio in SPCCI combustion while increasing the SI combustion ratio.

[0046] At high engine speed, a combustion cycle is shorter than at low engine speed. Since the amount of heat absorbed by an inner wall of the combustion chamber, etc., during combustion decreases accordingly, and the temperature within the combustion chamber decreases, it is difficult for the gas mixture to absorb heat, and a period for combustion of the gas mixture becomes short. Therefore, at high engine speed, SPCCI combustion must be stimulated to increase the SI combustion ratio under such an adverse condition. In this regard, by increasing the temperature of the gas mixture itself (the temperature immediately before combustion), SPCCI combustion is stimulated even under such a condition.

[0047] In SPCCI combustion, ignition of CI combustion occurs after SI combustion. Although increasing the gas mixture temperature stimulates both SI combustion and CI combustion, the combustion duration is short when the engine speed is high. Thus, SI combustion becomes sharper and is stimulated more than CI combustion, which starts later than SI combustion. Thus, increasing the gas mixture temperature increases the ratio of SI combustion at high engine speeds.

[0048] During engine operation, increasing the EGR ratio increases the ratio of EGR gas contained in the mixed gas (high-temperature combustion gas), thereby increasing the temperature of the mixed gas, and thus increasing the SI combustion ratio in SPCCI combustion. As a result, NVH is reduced below the allowable level even at high engine speeds. It should be noted that the EGR gas can be either internal EGR gas or external EGR gas, as long as it is at a high temperature.

[0049] When the engine speed exceeds a given limit start speed, the controller can output the control signal to the EGR system to limit the increase in the internal EGR ratio.

[0050] Although increasing the EGR ratio increases the temperature of the gas mixture, which suppresses NVH below the permissible level as described above, if the temperature of the gas mixture increases excessively, the CI combustion becomes harsh and combustion noise increases, which can lead to NVH exceeding the permissible level. Therefore, the controller limits the increase in the EGR ratio before such a situation can occur to prevent NVH from exceeding the permissible level.

[0051] In this case, the limit start speed can be set to be lower as the engine speed increases.

[0052] As engine load increases, the temperature of the exhaust gas mixture easily becomes excessively high. Setting the limit start speed, at which the EGR ratio increase limitation begins, lower with increasing engine load more reliably prevents NVH from exceeding the permissible level.

[0053] The engine or control system may further include an intake flow controller mounted on the engine and configured to change the flow of intake air introduced into the combustion chamber. The controller may output a control signal to the intake flow controller to increase the flow of intake air within an engine speed range that exceeds the limit starting speed.

[0054] Since the SI combustion ratio does not increase above the limit start speed, the effect of reducing combustion noise is limited and cannot be sufficiently maintained. Therefore, according to the above configuration, the effect limited by changing the EGR ratio is compensated. That is, by increasing the intake air flow, vaporization of injected fuel is stimulated, and SI combustion is performed in a state where the flow within the combustion chamber is strong. Thus, SI combustion becomes sharp and stimulated compared to CI combustion in which ignition is performed later. Therefore, the SI combustion ratio increases in SPCCI combustion, and NVH is more reliably prevented from exceeding the allowable value.

[0055] Furthermore, the controller may output a control signal to the spark plug so that an ignition timing advances within an engine speed range that exceeds the limit start speed.

[0056] By advancing the ignition timing, SI combustion starts early. Thus, SI combustion becomes sharp even in a short period of time at high engine speed, allowing the SI combustion ratio in SPCCI combustion to be increased. Therefore, by advancing the ignition timing, the effect limited by changing the EGR ratio is compensated, and NVH is reduced below the permissible level even at high engine speed.

[0057] The EGR gas may be an internal EGR gas.

[0058] Because the internal EGR gas is directly introduced into the combustion chamber, the high temperature is more easily maintained than with the external EGR gas, which is indirectly introduced into the combustion chamber. Therefore, the temperature of the gas mixture is slightly increased, and the SI combustion ratio in the SPCCI combustion is slightly increased.

[0059] In this case, the internal EGR gas can be introduced into the combustion chamber by providing a negative overlap period.

[0060] Alternatively, the internal EGR gas can be introduced into the combustion chamber by overlapping an open period of an intake valve with an open period of an exhaust valve (so-called positive overlap period).

[0061] When introducing the internal EGR gas by setting a negative overlap period (NVO setting), burned gas existing in the combustion chamber is used as the internal EGR gas as it is, the temperature of the gas mixture is slightly increased, and the ratio of SI combustion in the SPCCI combustion is slightly increased.

[0062] On the other hand, when the internal EGR gas is introduced by setting the positive overlap period (PVO setting), the high-temperature burned gas generated in the combustion chamber flows once from the combustion chamber to an intake tract. Then, the burned gas that has flowed into the intake tract is re-introduced into the combustion chamber. By flowing to the intake tract where the temperature is lower, the high-temperature burned gas introduced into the combustion chamber, i.e., the internal EGR gas, is cooled.

[0063] Therefore, the introduction of internal EGR gas through the PVO setting lowers the combustion chamber temperature more than the NVO setting. This slows down the CI combustion due to auto-ignition in the SPCCI combustion and prevents the increase in combustion noise due to excessive CI combustion.

[0064] Furthermore, in this case, the EGR ratio may continuously increase without restriction until the engine speed comes close to the highest speed within an operating range of the engine in which the combustion of the given mode is performed.

[0065] As described above, if the temperature of the gas mixture is excessively increased, since NVH may exceed the allowable value, the increase in the EGR ratio must be limited. In this regard, when the internal EGR gas is introduced through the NVO setting, the need to limit the increase in the EGR ratio is high because the temperature in the combustion chamber easily becomes excessively high. On the other hand, when the internal EGR gas is introduced through the PVO setting, the temperature in the combustion chamber is prevented from rising excessively. Thus, when the internal EGR gas is introduced through the PVO setting, the need to limit the increase in the EGR ratio is low, and the EGR ratio can be continuously high until the engine speed approaches the highest speed.

[0066] The engine or control system may further include a boosting system mounted on the engine and configured to boost gas to be introduced into the combustion chamber. The controller may output a control signal to the boosting system to perform boosting within a first range where an engine load is higher than a given load and not perform boosting within a second range where the engine load is below the given load. Within the second range, the controller may output the control signal to the EGR system to increase the EGR ratio higher at a high engine speed than at a low engine speed.

[0067] In such an engine, within the first range where boosting is performed, even if the internal EGR gas is introduced, the boost pressure will still scavenge. Therefore, within the first range, it is difficult to change the EGR ratio, and a state function within the combustion chamber becomes unstable, potentially causing deterioration in fuel efficiency. Therefore, within the second range where boosting is not performed, the controller outputs the control signal to the EGR system, causing the EGR ratio to increase. Thus, NVH is reduced below the allowable level even at high engine speeds.

[0068] A swirl flow may be formed inside the combustion chamber when combustion is carried out in the given mode.

[0069] When the swirl flow is formed in the combustion chamber, the flow of the gas mixture within the combustion chamber is stimulated and the gas mixture in the combustion chamber is also stratified. This prevents dangerous substances such as NO x and soot are generated, and the gas mixture that performs SI combustion in SPCCI combustion in a stable manner is formed unevenly around the spark plug at an ignition timing. This achieves more stable SPCCI combustion and reduces NVH below the permissible level even at high engine speeds.

[0070] According to a further aspect, a control or regulating device is provided for an internal combustion engine which can be operated in at least one given mode (SPCCI (SPark Controlled Compression Ignition) combustion) in which spark ignition combustion (SI combustion) and compression ignition combustion (CI combustion) are combined, in which, after a spark plug ignites the gas mixture to start the combustion, unburned gas mixture burns by self-ignition, wherein the control or regulating deviceThe control device comprises a controller having a heat ratio changing module configured to change, according to an operating state of the engine, a heat ratio as an index relating to a ratio of the heat amount generated by the SI combustion with respect to a total heat amount generated by the SPCCI combustion, and wherein the controller causes the heat ratio changing module to increase the heat ratio at a high engine speed compared to a low engine speed.

[0071] The control device may further include a pressure sensor configured to detect a pressure in the combustion chamber, wherein the controller is configured to set a target heat quantity ratio, receive a detection signal from the pressure sensor, calculate the heat quantity ratio based on a pressure waveform caused by the combustion of the gas mixture, and, when the calculated heat quantity ratio is different from the target heat quantity ratio, change the heat quantity ratio to approximate the target heat quantity ratio.

[0072] According to yet another aspect, there is provided a control device for an internal combustion engine that can be operated in at least one given mode in which, after a spark plug ignites the gas mixture to start combustion, unburned gas mixture burns by auto-ignition, and that includes an EGR system configured to change an EGR ratio, which is a ratio of an amount of EGR gas contained within the gas mixture within a combustion chamber, wherein the control device includes a controller operatively connected to the spark plug and the EGR system, and configured to operate the engine by outputting a control signal to the spark plug or the EGR system, wherein the controller outputs the control signal to the spark plug at a given ignition timing to perform combustion in the given mode, and the control signalOutputs a control signal to the EGR system to increase the EGR ratio higher at high engine speed than at low engine speed.

[0073] Preferably, the controller outputs the control signal to the EGR system to limit the increase in the EGR ratio when the engine speed exceeds a given limit start speed.

[0074] According to yet another aspect, a method for controlling an internal combustion engine is provided, comprising the steps of: Operating the engine in at least one given mode (SPCCI (SPark Controlled Compression Ignition) combustion) in which spark ignition combustion (SI combustion) and compression ignition combustion (CI combustion) are combined, in which, after a spark plug ignites the gas mixture to start combustion, unburned gas mixture burns by self-ignition, Changing, according to an operating condition of the engine, a heat quantity ratio as an index concerning a ratio of the heat quantity generated by the SI combustion with respect to a total heat quantity generated by the SPCCI combustion, and Increasing the heat quantity ratio at a high engine speed compared to a low engine speed.

[0075] According to yet another aspect, a method for controlling an internal combustion engine is provided, comprising the following steps: operating the engine in at least one given mode in which, after a spark plug ignites the gas mixture to start combustion, unburned gas mixture burns by auto-ignition, Changing an EGR ratio, which is a ratio of an amount of EGR gas contained in the gas mixture within a combustion chamber, Actuating the spark plug at a given ignition timing to perform combustion in the given mode, and Increasing the EGR ratio higher at a high engine speed than at a low engine speed.

[0076] According to a further aspect, a computer program product is provided comprising computer-readable instructions which, when loaded and executed on a suitable system, can perform the steps of any of the above-mentioned methods. BRIEF DESCRIPTION OF THE DRAWINGS Fig. Figure 1 is a diagram showing a configuration of an engine. Fig. 2 is a diagram showing a structure of a combustion chamber, in which the upper part is a plan view of the combustion chamber and the lower part is a II-II cross-sectional view. Fig. 3 is a plan view showing structures of the combustion chamber and an intake system. Fig. 4 is a block diagram showing a configuration of a control or Shows the engine control device. Fig. 5 is a diagram showing a test bench device for measuring a swirl ratio. Fig. 6 is a diagram showing a relationship between an opening ratio of a secondary passage and the swirl ratio. Fig. 7A is a graph showing a map of a first operating range. Fig. Figure 7B is a graph showing a map of a second operating range. Fig. Figure 7C is a graph showing a map of a third operating range. Fig. Figure 8 shows graphs conceptually showing a change in heat generation rate in SPCCI combustion in which SI combustion and CI combustion are combined. Fig. 9 is a graph showing a change in SI ratio, a change in a state function within a combustion chamber, a change in an overlap period between an intake valve and an exhaust valve, and changes in injection timing and ignition timing of fuel with respect to an engine load, corresponding to the first operating range map. Fig. 10 shows graphs in which the upper part shows a change in a combustion waveform due to an increase in engine load during non-boosted SPCCI combustion and the lower part shows a change in a combustion waveform according to an increase in engine load during boosted SPCCI combustion. Fig. 11 is a graph showing an example of a relationship between an engine speed and the SI ratio within an operating range in which the SPCCI combustion of the engine is performed according to the first operating range map. Fig. 12A and Fig. 12B are graphs showing an example of a relationship between the engine speed and an internal EGR ratio within the operating range in which the SPCCI combustion of the engine is performed according to the first operating range map, where Fig. 12A shows the relationship when the internal EGR gas is introduced by setting a negative overlap period, and Fig. 12B shows the relationship when the internal EGR gas is introduced by setting a positive overlap period. Fig. 13A and Fig. 13B are graphs showing an example of a relationship between the engine speed and an external EGR ratio within the operating range in which the SPCCI combustion of the engine is performed according to the first operating range map, where Fig. 13A shows the relationship when hardly any swirl flow is formed in the combustion chamber, and Fig. 13B shows the relationship when a swirl flow with a given strength is formed in the combustion chamber. Fig. 14A and Fig. 14B are graphs showing an example of a relationship between the engine speed and the opening of a swirl control valve within the operating range in which the SPCCI combustion of the engine is performed according to the first operating range map, wherein Fig. 14A shows the relationship in the engine according to the first operating range map and Fig. 14B shows the relationship in the engine according to a second or third operating range map. Fig. 15 is a graph showing an example of a relationship between the engine speed and the ignition timing within the operating range in which the SPCCI combustion of the engine is performed according to the first operating range map. Fig. 16 is a flowchart showing a flow of control of the engine executed by an ECU. Fig. Figure 17 is a diagram showing a control concept for adjusting the SI ratio. Fig. 18 is a diagram showing fuel injection timing, ignition timing, and combustion waveforms in respective operating states on the third operating range map. Fig. 19 is a short graphic of the Fig. 7C, which shows the respective operating states. Fig. 20 is a diagram showing combustion waveforms in respective operating conditions shown in Fig. 19 are shown. DETAILED DESCRIPTION OF REVELATION

[0077] Embodiments of a control system of a compression-ignition engine (hereinafter referred to simply as "engine 1") will be described in detail with reference to the accompanying drawings. The following description provides an example of the control system of engine 1. Fig. 1 is a diagram showing a configuration of engine 1. Fig. 2 is a cross-sectional view showing a structure of a combustion chamber. Fig. Figure 3 is a plan view showing the structures of the combustion chamber and an intake system. Note that in Fig. 1 is an inlet side on the left and an outlet side on the right side of the drawing sheet. Further, in Fig. 2 and Fig. 3 the inlet side on the right side and the outlet side on the left side of the drawing sheets. Fig. 4 is a block diagram showing a configuration of the control system of the engine 1.

[0078] Furthermore, the definition of "EGR gas" as used below includes "burned gas (exhaust gas) remaining in the combustion chamber and / or being drawn back into the combustion chamber." Similarly, the definition of "internal EGR gas" includes "burned gas (exhaust gas) remaining in the combustion chamber and / or being directly drawn back into the combustion chamber without passing through a passage external to the engine," and the definition of "external EGR gas" includes "burned gas (exhaust gas) indirectly drawn back into the combustion chamber via passages external to the engine, such as an exhaust passage and an intake passage." Furthermore, the "EGR ratio" is equivalent to "a ratio of an amount of EGR gas contained in a gas mixture (total gas) in the combustion chamber."In addition, the "internal EGR ratio" is equivalent to "a ratio of the amount of internal EGR gas contained in the mixed gas (total gas) in the combustion chamber," and the "external EGR ratio" is equivalent to "a ratio of the amount of external EGR gas contained in the mixed gas (total gas) in the combustion chamber." Details of the "heat quantity ratio (SI ratio)" will be described later. <spcci-verbrennung>

[0079] The engine 1 performs combustion in a mode in which SI combustion and CI combustion are combined.

[0080] SI combustion is a combustion-accompanying flame propagation process that begins when the gas mixture within a combustion chamber is forcibly ignited. CI combustion is a combustion process that begins when the gas mixture within the combustion chamber self-ignites through compression. In the combustion mode that combines SI combustion and CI combustion, a spark plug forcibly ignites the gas mixture within the combustion chamber to burn it through flame propagation. The heat generated by this combustion and pressure increase thereby increases the temperature in the combustion chamber, resulting in the combustion of unburned gas mixture through self-ignition.

[0081] In this regard, the change in the temperature within the combustion chamber before the start of compression can be reduced by adjusting the amount of heat generated during SI combustion. For example, by controlling the ignition timing to adjust the start time of SI combustion according to the temperature within the combustion chamber before the start of compression, the unburned gas mixture can self-ignite at a desired time.

[0082] Hereinafter, the combustion mode in which SI combustion and CI combustion are combined so that CI combustion is controlled using SI combustion is referred to as SPCCI combustion (Spark Controlled Compression Ignition (SPCCI) combustion). <Konfiguration des Motors 1>

[0083] The engine 1 is preferably mounted on a four-wheeled vehicle. The vehicle travels through the operation of the engine 1. The fuel of the engine 1 is gasoline in this embodiment. The gasoline may contain bioethanol, etc. The fuel of the engine 1 may be any type of fuel, as long as it is liquid fuel that contains at least gasoline.

[0084] The engine 1 includes a cylinder block 12 and a cylinder head 13 placed on the cylinder block 12. The cylinder block 12 is formed therein with a plurality of cylinders 11. In Fig. 1 and Fig. 2, only one cylinder 11 is shown. The engine 1 is preferably a multi-cylinder engine. The engine 1 also includes a piston 3, an injector 6, a spark plug 25, and at least one intake valve 21 and at least one exhaust valve 22. (Piston 3)

[0085] Piston 3 is reciprocally inserted into each cylinder 11. Piston 3 is connected to a crankshaft 15 via a connecting rod 14. Piston 3 defines a combustion chamber 17 together with cylinder 11 and cylinder head 13. Note that the definition of "combustion chamber" is not limited to a space formed when piston 3 is at top dead center on a compression stroke (CTDC), but can be broad. That is, "combustion chamber" can mean any space formed by piston 3, cylinder 11, and cylinder head 13, regardless of the position of piston 3.

[0086] As in Fig. As shown in FIG. 2, a lower surface of the cylinder head 13, i.e., a ceiling surface of the combustion chamber 17, is formed by an inclined surface 1311 and an inclined surface 1312. The inclined surface 1311 slopes toward an axis X2 (an axis passing through the injection center of the injector 6) from the intake side. The inclined surface 1312 slopes upward toward the axis X2 from the exhaust side. The ceiling surface of the combustion chamber 17 has a so-called pent-roof shape.

[0087] An upper surface of the piston 3 curves toward the ceiling surface of the combustion chamber 17 (or may be flat). A cavity 31 is formed in the upper surface of the piston 3. The cavity 31 is formed by notching the upper surface of the piston 3. The cavity 31 has a flat plate shape. The cavity 31 faces the injector 6 when the piston 3 is near CTDC.

[0088] The center of the cavity 31 is offset from an axis X1 (an axis passing through the radial center of the cylinder 11) toward the exhaust side and coincides with the axis X2. The cavity 31 has a convex portion 311. The convex portion 311 is formed along the axis X2. The convex portion 311 has a substantially conical shape. The convex portion 311 extends from the bottom of the cavity 31 upward to the ceiling surface of the cylinder 11.

[0089] The cavity 31 has a notched portion 312 formed to completely surround the convex portion 311. The cavity 31 has a symmetrical shape with respect to the axis X2. A squish area 171 is formed outside a portion of the cavity 31 within the combustion chamber 17.

[0090] A circumferential side surface of the notched portion 312 extends from a bottom surface of the cavity 31 to an opening surface of the cavity 31, which is inclined with respect to the axis X2. An inner diameter of the cavity 31 at the notched portion 312 gradually increases from the bottom surface of the cavity 31 to the opening surface of the cavity 31.

[0091] It should be noted that the shape of the combustion chamber 17 is not limited to the Fig. 2 is limited. The shapes of the cavity 31, the upper surface of the piston 3, the ceiling surface of the combustion chamber 17, etc., can be suitably changed.

[0092] For example, the depth of the notched portion 312 on the outer peripheral side may be smaller. In this case, the amount of EGR gas around the spark plug 25 decreases, and the flame propagation of the SI combustion in the SPCCI combustion becomes favorable.

[0093] As shown by a virtual line L1 from Fig. 2, the cavity 31 may also have an asymmetric shape with respect to the axis X2. That is, the notched portion 312 is radially larger and deeper on the exhaust side than on the intake side. In this case, a fuel concentration around the spark plug 25 increases at the time of ignition, and therefore, the ignitability of the SI combustion in the SPCCI combustion improves.

[0094] As shown by a virtual line L2 from Fig. 2, the cavity 31 does not need to have the convex portion 311. That is, the cavity 31 has a spherical shape that gradually flattens from the center radially outward. In this case, since it becomes more difficult for the piston 3 to come into contact with an intake valve 21 and an exhaust valve 22, the degree of freedom for controlling the opening and closing of the intake valve 21 and the exhaust valve 22 increases. In this case, when a swirl flow is formed in the combustion chamber 17, the flow stabilizes, and therefore stratification of the gas mixture becomes easy.

[0095] The geometric compression ratio of engine 1 is set to be between approximately 13:1 and approximately 30:1. As described above, SPCCI combustion controls CI combustion by utilizing the heat generated by SI combustion and the pressure increase. In this engine 1, it is not necessary to increase the temperature of combustion chamber 17 when piston 3 reaches CTDC for the gas mixture to self-ignite (i.e., the final compression temperature).

[0096] That is, even though engine 1 performs CI combustion, its geometric compression ratio can be set relatively low. Reducing the geometric compression ratio is beneficial for reducing cooling loss and mechanical loss. For example, the geometric compression ratio can be set to approximately 14:1 to approximately 17:1 in regular specifications (the fuel octane number is approximately 91) and approximately 15:1 to approximately 18:1 in high-octane specifications (the fuel octane number is approximately 96). (Inlet valve 21 and exhaust valve 22)

[0097] The cylinder head 13 is formed with an inlet port 18 for each cylinder 11. As in Fig. 3, the intake port 18 includes two intake ports, a first intake port 181 and a second intake port 182. The first intake port 181 and the second intake port 182 are arranged in axial directions of the crankshaft 15, that is, in the forward and reverse directions of the engine 1. The intake port 18 communicates with the combustion chamber 17. Although not shown in detail, the intake tract 18 is a so-called tumble port. That is, the intake port 18 has a shape such that a tumble flow is formed in the combustion chamber 17.

[0098] The intake valve 21 is arranged in the intake port 18. The intake valve 21 opens and closes the intake port 18 to and from the combustion chamber 17. The intake valve 21 is opened and closed by a valve actuating mechanism at a given time. This valve actuating mechanism may be a variable valve mechanism that makes a valve timing and / or a valve lift variable. In this configuration example, the variable valve mechanism, as shown in Fig. 4, an electrically operated intake S-VT (Sequential Valve Timing) 23. The electrically operated intake S-VT 23 continuously varies a rotational phase of an intake camshaft within a given angular range. Thus, the opening and closing timing of the intake valve 21 continuously changes. Note that the intake valve actuation mechanism may include a hydraulically operated S-VT instead of the electrically operated S-VT.

[0099] The cylinder head 13 is also formed with an exhaust port 19 for each cylinder 11. As in Fig. 3, the exhaust port 19 also includes two exhaust ports, a first exhaust port 191 and a second exhaust port 192. The first exhaust port 191 and the second exhaust port 192 are arranged in the front-rear direction of the engine 1. The exhaust port 19 communicates with the combustion chamber 17.

[0100] The exhaust valve 22 is arranged in the exhaust port 19. The exhaust valve 22 opens and closes the exhaust port 19 to and from the combustion chamber 17. The exhaust valve 22 is opened and closed at a given time by a valve actuation mechanism. This valve actuation mechanism may be a variable valve mechanism that makes a valve timing and / or a valve lift variable. In this configuration example, as shown in Fig. As shown in Figure 4, the variable valve mechanism includes an electrically operated exhaust S-VT 24. The electrically operated exhaust S-VT 24 continuously varies the rotational phase of an exhaust camshaft within a given angular range. Thus, the opening and closing timing of the exhaust valve 22 continuously changes. Note that the exhaust valve actuation mechanism may include a hydraulically operated S-VT instead of the electrically operated S-VT.

[0101] Although it will be described in detail later, the engine 1 adjusts the length of an overlap period of the opening timing of the intake valve 21 and the closing timing of the exhaust valve 22 by the electrically operated intake S-VT 23 and the electrically operated exhaust S-VT 24. Thus, residual gas in the combustion chamber 17 is purged. Further, by adjusting the length of the overlap period, internal EGR gas is introduced into the combustion chamber 17 or is trapped in the combustion chamber 17. In this configuration example, the electrically operated intake S-VT 23 and the electrically operated exhaust S-VT 24 constitute an internal EGR system as one of the state function setting devices. Note that the internal EGR system is not necessarily constituted by the S-VT. (Injector 6)

[0102] The injector 6 is mounted on the cylinder head 13 for each cylinder 11. The injector 6 injects the fuel directly into the combustion chamber 17. The injector 6 is arranged in a valley section of the pent roof where the inclined surface 1311 on the intake side and the inclined surface 1312 on the exhaust side intersect. As shown in Fig. 2, the injector 6 is arranged so that its injection axis extends substantially along the axis X2. The injection axis of the injector 6 coincides with the position of the convex portion 311 of the cavity 31. The injector 6 is oriented toward the cavity 31. It should be noted that the injection axis of the injector 6 may coincide with the central axis X1 of the cylinder. In this case, too, it is desirable for the injection axis of the injector 6 to coincide with the position of the convex portion 311 of the cavity 31.

[0103] Although not shown in detail, the injector 6 is formed by a multi-port fuel injector having a plurality of nozzle openings. As indicated by the two-dot chain lines in Fig. As indicated in Figure 2, the injector 6 injects the fuel such that the fuel spray spreads radially downward from an upper central portion of the combustion chamber 17, where the injection center is located. In this configuration example, the injector 6 has ten nozzle openings, and the nozzle openings are arranged at a uniform angular pitch in the circumferential direction.

[0104] As in the upper part of Fig. 2, the axes of the nozzle openings do not circumferentially overlap with the spark plug 25 described later. That is, the spark plug 25 is sandwiched between the axes of two adjacent nozzle openings. This prevents the fuel spray injected from the injector 6 from directly impinging on the spark plug 25 and wetting an electrode.

[0105] As described later, the injector 6 may inject the fuel at the time when the piston 3 is positioned near the CTDC. In this case, when the injector 6 injects the fuel, the fuel spray flows downward along the convex portion 311 of the cavity 31 while mixing with fresh air, and flows along the bottom surface and the peripheral surface of the notched portion 312 to disperse radially outward from the center of the combustion chamber 17. Then, the mixed gas reaches the opening of the cavity 31, flows along the inclined surface 1311 on the intake side and the inclined surface 1312 on the exhaust side, and further flows from the outer peripheral side toward the center of the combustion chamber 17. Note that the injector 6 is not limited to the multi-port injector. The injector 6 may employ an outward-opening valve injector.

[0106] A fuel supply system 61 is connected to the injector 6. The fuel supply system 61 includes a fuel tank 63 configured to store the fuel and a fuel supply path 62 connecting the fuel tank 63 to the injector 6. A fuel pump 65 and a common rail 64 are provided in the fuel supply path 62. The fuel pump 65 pumps the fuel to the common rail 64. In this embodiment, the fuel pump 65 is a piston pump driven by the crankshaft 15.

[0107] The common rail 64 stores the fuel pumped at high fuel pressure by the fuel pump 65. When the injector 6 opens, the fuel stored in the common rail 64 is injected into the combustion chamber 17 from the nozzle holes of the injector 6. The fuel supply system 61 supplies the fuel to the injector 6 at a high pressure of about 30 MPa or more. For example, a maximum fuel pressure of the fuel supply system 61 may be about 120 MPa. The pressure of the fuel supplied to the injector 6 can be changed according to an operating state of the engine 1. Note that the structure of the fuel supply system 61 is not limited to the above structure. (Spark plug 25)

[0108] The spark plug 25 is mounted on the cylinder head 13 for each cylinder 11. The spark plug 25 forcibly ignites the gas mixture in the combustion chamber 17. In this configuration example, the spark plug 25 is arranged on the intake side of the center axis X1 of the cylinder 11. The spark plug 25 is located between the two intake ports 18. The spark plug 25 is mounted on the cylinder head 13 so as to extend substantially downward toward the center of the combustion chamber 17 in an inclined position with respect to the up-down directions of the cylinder head 13.

[0109] As in Fig. 2, the electrode of the spark plug 25 is located near the ceiling surface of the combustion chamber 17 so as to be directed toward the combustion chamber 17. The electrode of the spark plug 25 is adjacent to the nozzle openings of the injector 6. It should be noted that the arrangement position of the spark plug 25 is not limited to the configuration example of Fig. 2 is limited. The spark plug 25 may be arranged on the exhaust side of the axis X1. Alternatively, the spark plug 25 may be arranged on the axis X1, and the injector 6 may be arranged on the intake side or the exhaust side with respect to the axis X1. (Inlet tract 40)

[0110] An intake tract 40 is connected to one side of the engine 1. The intake tract 40 communicates with the intake ports 18 of the cylinders 11. The intake tract 40 is a tract through which gas to be introduced into the combustion chamber 17 flows. An air cleaner 41 that filters fresh air is arranged in an upstream end portion of the intake tract 40. An expansion tank 42 is arranged near a downstream end of the intake tract 40. Although not shown in detail, a portion of the intake tract 40 downstream of the expansion tank 42 forms independent tracts branched for the respective cylinders 11. Downstream ends of the independent tracts are respectively communicated with the intake ports 18 of the cylinders 11.

[0111] A throttle valve 43 is arranged in the intake tract 40 between the air cleaner 41 and the surge tank 42. The throttle valve 43 adjusts the amount of fresh air introduced into the combustion chamber 17 by changing its opening. The throttle valve 43 constitutes one of the state function adjustment devices.

[0112] A booster 44 is arranged in the intake tract 40 downstream of the throttle valve 43. The booster 44 boosts the gas introduced into the combustion chamber 17. In this configuration example, the booster 44 is an engine-driven supercharger. The booster 44 may be, for example, a Lisholm type. The booster 44 may have any structure. The booster 44 may be a Roots type, a vane type, or a centrifugal type. Note that the booster may be an electric booster or a turbocharger driven by the exhaust gas.

[0113] An electromagnetic clutch 45 is preferably arranged between the booster 44 and the engine 1. The electromagnetic clutch 45 controls the flow of driving force between the booster 44 and the engine 1, for example, transmitting the driving force from the engine 1 to the booster 44 or interrupting the transmission of the driving force therebetween. As described later, the on / off of the booster 44 is switched by an ECU 10 (Engine Control Unit) that switches the connection / disconnection of the electromagnetic clutch 45. That is, in this engine 1, switching can be carried out between boosting gas to be introduced into the combustion chamber 17 by the booster 44 and non-boosting gas to be introduced into the combustion chamber 17 by the booster 44.

[0114] An intercooler 46 is arranged in the intake tract 40 downstream of the booster 44. The intercooler 46 cools the gas compressed in the booster 44. The intercooler 46 can, for example, be a water-cooled type.

[0115] A bypass tract 47 is connected to the intake tract 40. The bypass tract 47 connects a portion of the intake tract 40 upstream of the booster 44 with a portion of the intake tract 40 downstream of the charge air cooler 46, in particular the expansion tank 42. An air bypass valve 48 is arranged in the bypass tract 47. The air bypass valve 48 changes a flow rate of the gas flowing through the bypass tract 47.

[0116] When the booster 44 is off (i.e., when the electromagnetic clutch 45 is disengaged), the air bypass valve 48 is substantially fully open. Thus, the gas flowing through the intake tract 40 bypasses the booster 44 and is introduced into the combustion chamber 17 of the engine 1. The engine 1 operates in a non-boosted state, i.e., in a naturally aspirated state.

[0117] When the booster 44 is turned on (that is, when the electromagnetic clutch 45 is connected), it operates in a boost state (a state in which a pressure higher than atmospheric pressure is dynamically applied to the downstream side of the booster 44). The gas passed through the booster 44 partially flows back upstream of the booster 44 through the bypass passage 47. By controlling an opening of the air bypass valve 48, a backflow amount is changed, resulting in adjusting the boost pressure of the gas introduced into the combustion chamber 17. In this configuration example, a boosting system 49 includes the booster 44, the bypass passage 47, and the air bypass valve 48. The air bypass valve 48 constitutes one of the state function adjusting devices. (Swirl control or regulating valve 56)

[0118] Further, a swirl control valve (SCV) 56 (intake flow control device) is arranged in the intake tract 40, which controls a flow of intake air introduced into the combustion chamber 17 to form a swirl flow therein and changes the strength of the flow.

[0119] As in Fig. As shown in Figure 3, the SCV 56 is disposed in a secondary passage 402. The secondary passage 402 is one of a primary passage 401 and a secondary passage 402, which communicate with the first intake port 181 and the second intake port 182, respectively. The SCV 56 is an opening control valve capable of adjusting the opening of a cross-section of the secondary passage.

[0120] When the opening of the SCV 56 is small, the flow rate of intake air into the combustion chamber 17 from the first intake port 181 relatively increases, while the flow rate of intake air into the combustion chamber 17 from the second intake port 182 relatively decreases. Thus, the swirl flow in the combustion chamber 17 becomes strong.

[0121] When the opening of the SCV 56 is large, the flow rates of intake air into the combustion chamber 17 from the first intake port 181 and the second intake port 182 become substantially uniform, and thus the swirl flow in the combustion chamber 17 becomes weak. When the SCV 56 is substantially fully opened, no swirl flow occurs. It should be noted that the swirl flow is in the counterclockwise direction of Fig. 3 circulates, as indicated by the white-bordered arrows (see also the white-bordered arrows of Fig. 2).

[0122] Note that, alternatively / in addition to attaching the SCV 56 to the intake tract 40, a structure may be used in which the opening periods of the two intake valves 21 are varied to introduce the intake air into the combustion chamber 17 from only one of the intake valves 21. By opening only one of the two intake valves 21, the intake air is unevenly introduced into the combustion chamber 17, thus generating the swirl flow in the combustion chamber 17. Alternatively, the shapes of the intake ports 18 may be designed to generate the swirl flow in the combustion chamber 17. (swirling flow)

[0123] Here, the strength of the swirl flow in the combustion chamber 17 is defined. In this configuration example, the strength of the swirl flow in the combustion chamber 17 is expressed by a "swirl ratio." The "swirl ratio" can be defined as a value obtained by dividing a value obtained by measuring a lateral angular velocity of the intake flow for each valve lift and integrating the value by an engine angular velocity. The lateral angular velocity of the intake flow can be determined based on a measurement using a Fig. 5 shown test bench device.

[0124] In particular, the Fig. 5, the cylinder head 13 is placed upside down on a pedestal to connect the intake port 18 to an intake air supply device (not shown), and a cylinder 36 is placed on the cylinder head 13 to connect at its upper end to a momentum meter 38 having a honeycomb rotor 37. A lower surface of the momentum meter 38 is positioned at a position 1.75D (where "D" is a cylinder bore diameter) away from a mating surface between the cylinder head 13 and the cylinder 36. The momentum meter 38 measures a torque acting on the honeycomb rotor 37 by a swirl generated in the cylinder 36 according to the supply of intake air (see the arrow of Fig. 5), and the lateral angular velocity of the inlet flow is obtained based on the torque.

[0125] Fig. Figure 6 illustrates a relationship between the opening of the SCV 56 of the engine 1 and the swirl ratio. In Fig. 6, the opening of the SCV 56 is expressed by an opening ratio relative to the cross-section of the secondary passage 402 when it is fully open. The opening ratio of the secondary passage 402 is approximately 0% when the SCV 56 is fully closed and increases from 0% as the opening of the SCV 56 increases. The opening ratio of the secondary passage 402 is approximately 100% when the SCV 56 is fully open.

[0126] As in Fig. 6, in the engine 1, the swirl ratio becomes approximately 6 when the SCV 56 is fully closed. To set the swirl ratio to 4 or higher, the opening of the SCV 56 can be adjusted within a range of the opening ratio from approximately 0 to approximately 15%.

[0127] To set the swirl ratio to less than 4, the opening of the SCV 56 can be set within an opening ratio range below approximately 15%. Specifically, to ensure the fluidity of the gas mixture and to control the stratification of the gas mixture in the combustion chamber 17 in conjunction with SPCCI combustion within the operating range where the engine load is low or medium, the swirl ratio is preferably set within a range of approximately 1.5 to approximately 3 (approximately 25% to approximately 40% relative to the opening of the SCV 56). (Outlet tract 50)

[0128] An exhaust duct 50 is connected to a side of the engine 1 opposite the intake duct 40. The exhaust duct 50 communicates with the exhaust ports 19 of the cylinders 11. The exhaust duct 50 is a duct through which the exhaust gas discharged from the combustion chamber 17 flows. Although not shown in detail, an upstream portion of the exhaust duct 50 forms independent ducts branched for the respective cylinders 11. Upstream ends of the independent ducts are respectively connected to the exhaust ports 19 of the cylinders 11.

[0129] An exhaust gas purification system with one or more catalysts is arranged in the exhaust tract 50. The exhaust gas purification system of this configuration example has two catalysts. The upstream-side catalyst is arranged in an engine compartment and includes a three-way catalyst 511 and a GPF (gasoline particulate filter) 512. The downstream-side catalyst is arranged outside the engine compartment and includes a three-way catalyst.

[0130] Note that the exhaust gas purification system is not limited to the illustrated structure. For example, the GPF 512 may be omitted. Furthermore, the catalyst is not limited to the three-way catalyst. Furthermore, the arrangement order of the three-way catalyst and the GPF may be appropriately changed.

[0131] An EGR passage 52, which forms an external EGR system, is connected between the intake passage 40 and the exhaust passage 50. The EGR passage 52 is a passage for returning a portion of the burned gas to the intake passage 40. An upstream end of the EGR passage 52 is connected to the exhaust passage 50 between the upstream catalyst and the downstream catalyst. A downstream end of the EGR passage 52 is connected to the intake passage 40 upstream of the booster 44.

[0132] For example, the downstream end of the EGR tract 52 is connected to an intermediate position of the bypass tract 47. The EGR gas flowing through the EGR tract 52 enters the intake tract 40 upstream of the booster 44 without passing through the air bypass valve 48 of the bypass tract 47.

[0133] A water-cooled EGR cooler 53 is disposed in the EGR passage 52. The EGR cooler 53 cools the burned gas. An EGR valve 54 is also disposed in the EGR passage 52. The EGR valve 54 adjusts the flow rate of the burned gas in the EGR passage 52. By changing an opening of the EGR valve 54, the recirculation amount of the cooled burned gas (i.e., external EGR gas) is changed.

[0134] In this configuration example, an EGR system 55 includes an external EGR system including the EGR tract 52 and the EGR valve 54, and an internal EGR system including the electrically operated intake S-VT 23 and the electrically operated exhaust S-VT 24 described above. The EGR valve 54 constitutes one of the state function setting devices. In the external EGR system, since the EGR tract 52 is connected downstream of the upstream catalyst and the EGR cooler 53 is provided, the burned gas is supplied to the combustion chamber 17 at a lower temperature than in the internal EGR system. (ECU 10)

[0135] A control system 20 of the engine 1 includes an ECU (Engine Control Unit) 10 configured to operate the engine 1. The ECU 10 is a controller based on a well-known microcomputer. The ECU 10 includes a central processing unit (CPU) 101 comprising a processor configured to execute program(s) / instructions, a memory 102 consisting of RAM(s) (Random Access Memory) and ROM(s) (Read Only Memory) and configured to store the program(s) / instructions and data, and an input / output bus 103 configured to input and output electrical signals.

[0136] The memory 102 stores an SI ratio change module (heat quantity ratio change module) 102a, which changes the SI ratio (heat quantity ratio, described in detail later) according to the operating state of the engine 1 to control the SPCCI combustion. The SI ratio change module 102a includes, for example, data such as a control program and maps used for the control program.

[0137] As in Fig. 1 and Fig. As shown in Figure 4, various sensors SW1 to SW16 are connected to the ECU 10. The sensors SW1 to SW16 output detection signals to the ECU 10. The sensors include the following sensors.

[0138] That is, the sensors include an air flow sensor SW1, which is arranged in the intake tract 40 downstream of the air cleaner 41 and is configured to detect the flow rate of fresh air in the intake tract 40, a first intake air temperature sensor SW2, which is also arranged in the intake tract 40 downstream of the air cleaner 41 and is configured to detect a temperature of the fresh air, a first pressure sensor SW3, which is arranged in the intake tract 40 downstream of the connection position with the EGR tract 52 and upstream of the booster 44 and is configured to detect pressure of the gas flowing into the booster 44, a second intake air temperature sensor SW4, which is arranged in the intake tract 40 downstream of the booster 44 and upstream of the connection position of the bypass tract 47 and is configured to detect a temperature of the gas flowing out of the booster 44, a second Pressure sensor SW5,which is attached to the expansion tank 42 and is configured to detect pressure of the gas at a position downstream of the booster 44, pressure sensors SW6 which are attached to the cylinder head 13 corresponding to the cylinders 11 and are configured to detect pressure in the combustion chambers 17, an exhaust gas temperature sensor SW7 which is arranged in the exhaust tract 50 and is configured to detect a temperature of the exhaust gas discharged from the combustion chamber 17, a linear O2 sensor SW8 which is arranged in the exhaust tract 50 upstream of the catalyst 511 and is configured to detect an oxygen concentration within the exhaust gas, a lambda O2 sensor SW9 which is arranged in the exhaust tract 50 downstream of the catalyst 511 and is configured to detect an oxygen concentration within the exhaust gas, a water temperature sensor SW10 which is attached to the engine 1 and is configured,to detect a temperature of the cooling water, a crank angle sensor SW11 mounted on the engine 1 and configured to detect a rotation angle of the crankshaft 15, an accelerator opening or position sensor SW12 mounted on an accelerator pedal mechanism and configured to detect an accelerator opening or position corresponding to an operation amount of an accelerator pedal, an intake cam angle sensor SW13 mounted on the engine 1 and configured to detect a rotation angle of the intake camshaft, an exhaust cam angle sensor SW14 mounted on the engine 1 and configured to detect a rotation angle of the exhaust camshaft, an EGR pressure difference sensor SW15 arranged in the EGR tract 52 and configured to detect a pressure difference between positions upstream and downstream of the EGR valve 54, and a fuel pressure sensor SW16,which is attached to the common rail 64 of the fuel supply system 61 and is configured to detect the pressure of the fuel to be supplied to the injector 6.,

[0139] Based on these detection signals, the ECU 10 determines the operating state of the engine 1 and calculates control amounts of the various devices. The ECU 10 outputs control signals related to the calculated control amounts to the injector 6, the spark plug 25, the electrically operated intake S-VT 23, the electrically operated exhaust S-VT 24, the fuel supply system 61, the throttle valve 43, the EGR valve 54, the electromagnetic clutch 45 of the booster 44, the air bypass valve 48, and the SCV 56.

[0140] For example, the ECU 10 changes the boost pressure by changing the opening of the air bypass valve 48 based on a pressure difference between the upstream and downstream sides of the booster 44 obtained from the detection signals of the first pressure sensor SW3 and the second pressure sensor SW5. Furthermore, the ECU 10 changes the amount of external EGR gas introduced into the combustion chamber 17 by changing the opening of the EGR valve 54 based on the pressure difference between the upstream and downstream sides of the EGR valve 54 obtained from the detection signal of the EGR pressure difference sensor SW15. Details of other controls of the engine 1 by the ECU 10 will be described later. (Engine operating range)

[0141] Fig. 7A shows a first configuration example of an operating range map of the engine 1 (first operating range map 700). The operating range map 700 is defined by an engine load and an engine speed and is roughly divided into the following four ranges based on the engine load and engine speed. (A): a low-load range including idle operation (B): a medium load range between the low load range (A) and the following high load range (C) (C): a high-load range including full engine load (D): a high-speed range (D) in which the engine speed is higher than in the low-load range (A), the medium-load range (B) and the high-load range (C).

[0142] Engine 1 executes SPCCI combustion within the medium load range (B) to improve fuel efficiency and exhaust emission performance. The combustion modes in each of the low load range (A), medium load range (B), and high load range (C) are described in detail below. (low load range)

[0143] When the operating state of the engine 1 is within the low-load range (A), the fuel injection amount is small. Therefore, the amount of heat generated when the gas mixture is burned in the combustion chamber 17 is small, and the temperature of the combustion chamber 17 is low. Since the temperature of the exhaust gas is also low, even if the internal EGR gas is introduced into the combustion chamber 17, as described later, it is difficult to raise the temperature of the combustion chamber 17 to such an extent that self-ignition is performed stably.

[0144] The combustion mode when the operating condition of the engine 1 is within the low-load range (A) is SI combustion, in which the spark plug 25 ignites the gas mixture within the combustion chamber 17 to burn it through flame propagation. Hereinafter, the combustion mode within the low-load range (A) may be referred to as "low-load SI combustion."

[0145] When the operating condition of engine 1 is within the low-load range (A), the air-fuel ratio (A / F) of the gas mixture is substantially at the theoretical air-fuel ratio (A / F = approximately 14.7:1). Note that, hereinafter, the air-fuel ratio, excess air ratio λ, and the G / F (gas / fuel ratio) of the gas mixture are the values ​​taken at the ignition timing.

[0146] When the air-fuel ratio of the gas mixture is adjusted substantially to the theoretical air-fuel ratio, the three-way catalyst is able to purify the exhaust gas discharged from the combustion chamber 17, thus improving the exhaust performance of the engine 1. The A / F of the gas mixture can be adjusted to remain within the purification window of the three-way catalyst (i.e., an air-fuel ratio width exhibiting the three-way purification function). Therefore, the excess air ratio λ of the gas mixture can be set to about 1.0 ± about 0.2.

[0147] To improve the fuel efficiency of the engine 1, when the operating condition of the engine 1 is within the low-load range (A), the EGR system 55 introduces the EGR gas into the combustion chamber 17. The G / F of the gas mixture, which is a mass ratio of the total gas to the fuel in the combustion chamber 17, is set between about 18:1 and about 30:1. The gas mixture is EGR-lean and has a high dilution ratio.

[0148] By setting the G / F of the gas mixture to, for example, approximately 25:1 within the low-load range (A), SI combustion is performed stably without the gas mixture self-igniting. Within the low-load range (A), the G / F of the gas mixture is kept constant regardless of the engine load. Thus, SI combustion is stable throughout the entire low-load range. Additionally, the fuel efficiency of engine 1 and the exhaust gas performance are improved.

[0149] When the operating condition of the engine 1 is within the low-load range (A), since the fuel quantity is low, a charge amount of gas into the combustion chamber 17 must be less than about 100% to bring λ of the gas mixture to about 1.0 ± about 0.2 and G / F to a value between about 18:1 and about 50:1. For example, the engine 1 performs throttling to change the opening of the throttle valve 43 and / or a mirror cycle to retard the closing timing of the intake valve 21 to after a bottom dead center (BDC) in the intake stroke.

[0150] When the engine 1 adopts the operating range map 700 when the operating condition is within the low-load range (A), the SCV 56 is substantially fully open. Therefore, the swirl flow hardly occurs in the combustion chamber 17.

[0151] It should be noted that within a low-load and low-speed segment of the low-load range (A), the combustion temperature of the gas mixture and the exhaust gas temperature can be increased by making the charge amount of the gas even smaller. This is advantageous for maintaining the catalyst 51 in an active state. (Medium load range)

[0152] When the operating condition of engine 1 is within the medium-load range (B), the fuel injection amount increases. The temperature of combustion chamber 17 rises, and thus, auto-ignition is performed in a stable manner. In the medium-load range (B), the engine performs CI combustion to improve fuel efficiency and exhaust emission performance.

[0153] In combustion caused by auto-ignition, the timing of auto-ignition changes significantly as the temperature within the combustion chamber varies before compression begins. Therefore, SPCCI combustion is performed within the medium-load range (B).

[0154] Furthermore, within the medium load range (B), the engine 1 adjusts the condition within the combustion chamber 17 so that λ of the gas mixture becomes about 1.0 ± about 0.2 and the G / F of the gas mixture becomes a value between about 18:1 and about 50:1. In addition, at the ignition timing, a required temperature T IG within the combustion chamber 17 about 570 to about 800 K, a required pressure P IG within the combustion chamber 17 is about 400 to about 920 kPa and kinetic turbulence energy within the combustion chamber 17 is about 17 to about 40 m 2 / s 2 .

[0155] When the engine 1 assumes the operating range map 700 when the operating condition is within the medium load range (B), the SCV 56 is substantially fully open. Therefore, the swirl flow in the combustion chamber 17 hardly occurs.

[0156] By precisely controlling the auto-ignition timing, an increase in combustion noise is avoided when the operating condition of engine 1 is within the mid-load range (B). Furthermore, by increasing the dilution ratio of the gas mixture as high as possible and performing CI combustion, the fuel efficiency of engine 1 is improved. By setting the mid-load range of the gas mixture to approximately 1.0 ± approximately 0.2, the three-way catalyst is able to purify the exhaust gas, thus improving the exhaust performance of engine 1.

[0157] As described above, within the low-load range (A), the G / F of the gas mixture is set between approximately 18:1 and approximately 50:1 (e.g., approximately 25:1), and λ of the gas mixture is set to approximately 1.0 ± approximately 0.2. The state function within the combustion chamber 17 does not vary greatly between the states in which the operating state of the engine 1 is within the low-load range (A) and within the medium-load range (B). Therefore, the robustness of the control of the engine 1 to changes in the engine load is improved.

[0158] When the operating state of the engine 1 is within the medium load range (B), unlike when it is within the low load range (A), the fuel quantity increases, so the charge quantity of the gas introduced into the combustion chamber 17 does not need to be adjusted. The throttle valve 43 is essentially fully open.

[0159] As the engine load increases and the fuel quantity further increases, in the suction state, the introduction amount of gas into the combustion chamber 17 becomes insufficient to adjust λ of the gas mixture to about 1.0 ± about 0.2 and the G / F of the gas mixture between about 18:1 and about 50:1. Therefore, in a segment of the middle load range (B) where the engine load is higher than a given load, the booster 44 boosts the gas to be introduced into the combustion chamber 17.

[0160] The medium-load range (B) is divided into a first medium-load segment (B1), in which the engine load is higher than the given load and boosting is performed, and a second medium-load segment (B2), in which the engine load is lower than the given load and boosting is not performed. The given load is, for example, approximately 1 / 2 the load. The second medium-load segment (B2) is a segment in which the engine load is lower than in the first medium-load segment (B1). Hereinafter, the combustion mode within the first medium-load segment (B1) may be referred to as "boosted SPCCI combustion," and the combustion mode within the second medium-load segment (B2) may be referred to as "unboosted SPCCI combustion."

[0161] Within the second mid-load segment (B2), in which boosting is not performed, the introduction amount of fresh air into the combustion chamber 17 increases, while the EGR gas decreases as the fuel amount increases. The G / F of the gas mixture decreases as the engine load increases. Since the throttle valve 43 is substantially fully open, the engine 1 changes the introduction amount of EGR gas into the combustion chamber 17 to change the amount of fresh air introduced into the combustion chamber 17. Within the second mid-load segment (B2), the state function within the combustion chamber 17 is adjusted so that, for example, λ of the gas mixture is substantially constant at about 1.0, while the G / F of the gas mixture is changed between about 25:1 and about 28:1.

[0162] On the other hand, within the first mid-load segment (B1) where boosting is performed, the engine 1 increases the introduction amounts of fresh air and EGR gas into the combustion chamber 17 as the fuel quantity increases. Thus, the G / F of the gas mixture is constant even as the engine load increases. For example, in the state function within the combustion chamber 17 within the first mid-load segment (B1), λ of the gas mixture is substantially constant at about 1.0, and the G / F of the gas mixture is constant at about 25:1. (high load range)

[0163] The combustion mode when the operating condition of engine 1 is within the high-load range (C) is SI combustion. This prioritizes the prevention of combustion noise. Hereinafter, the combustion mode within the high-load range may be referred to as "high-load SI combustion."

[0164] When the operating condition of the engine 1 is within the high-load range (C), λ of the gas mixture becomes about 1.0 ± about 0.2, and the G / F of the gas mixture is substantially adjusted between about 18:1 and about 30:1. The G / F of the gas mixture can be adjusted between about 18:1 and about 50:1. Within the high-load range (C), the throttle valve 43 is substantially fully open, and the booster 44 performs boosting.

[0165] In the high-load range (C), the engine 1 reduces the amount of EGR gas as the engine load increases. The G / F of the gas mixture decreases as the engine load increases. The amount of fresh air introduced into the combustion chamber 17 increases by the reduced amount of EGR gas, thus allowing the fuel amount to be increased, which is beneficial for increasing the maximum power of the engine 1.

[0166] In the engine 1 adopting the operating range map 700, when the operating condition is within the high-load range (C), the SCV 56 is substantially fully open. Therefore, swirl flow hardly occurs in the combustion chamber 17.

[0167] The state function within the combustion chamber 17 does not vary greatly between the states in which the operating state of the engine 1 is within the high-load range (C) and within the medium-load range (B). Therefore, the robustness of the control or regulation of the engine 1 with respect to changes in the engine load is improved.

[0168] Since the engine 1 performs the SI combustion within the high load range (C) as described above, there is a problem that abnormal combustion such as preignition and knocking easily occurs.

[0169] Therefore, in the high-load range (C), abnormal combustion in the engine 1 is avoided by developing the fuel injection mode. For example, the ECU 10 outputs control signals to the fuel supply system 61 and the injector 6 to inject fuel into the combustion chamber 17 at a high fuel pressure of approximately 30 MPa or higher at a time point in a period from a final stage of the compression stroke to an early stage of the expansion stroke (hereinafter, this period is referred to as a "lag period"). The ECU 10 also outputs a control signal to the spark plug 25 to ignite the gas mixture at a time point near CTDC after the fuel injection. Hereinafter, the fuel injection into the combustion chamber 17 at the high fuel pressure at the time point in the lag period is referred to as "high-pressure lag injection."

[0170] High-pressure delay injection shortens the reaction time of the gas mixture to prevent abnormal combustion. That is, the reaction time of the gas mixture is a total length of (1) a period for which the injector 6 injects fuel (i.e., injection period), (2) a period for which a combustible gas mixture forms around the spark plug 25 after fuel injection (i.e., gas mixture formation period), and (3) a period from the start of ignition to the end of SI combustion (i.e., combustion period).

[0171] The injection period and the gas mixture formation period become shorter by injecting fuel into the combustion chamber 17 at high fuel pressure. By shortening the injection period and the gas mixture formation period, the fuel injection start timing approaches the ignition timing. Since, in high-pressure delay injection, fuel injection into the combustion chamber 17 is performed at high fuel pressure, the fuel is injected at a time within the delay period from the final stage of the compression stroke to the early stage of the expansion stroke.

[0172] Injecting fuel into the combustion chamber 17 at high fuel pressure increases the kinetic turbulence energy within the combustion chamber 17. By bringing the fuel injection timing close to CTDC, it is possible to start SI combustion while the kinetic turbulence energy within the combustion chamber 17 is high. As a result, the combustion period becomes short.

[0173] Thus, in high-pressure delay injection, since the injection period, the gas mixture formation period, and the combustion period are each shortened, the reaction time of the gas mixture is significantly shortened compared to a case where the fuel is injected into the injection chamber 17 during the intake stroke. As a result, abnormal combustion is avoided.

[0174] In the field of engine control, traditionally, the ignition timing is retarded to prevent abnormal combustion. However, retardation of the ignition timing degrades fuel efficiency. With high-pressure delay injection, the ignition timing does not need to be retarded. Therefore, fuel efficiency is improved by using high-pressure delay injection.

[0175] By setting the fuel pressure to, for example, approximately 30 MPa or higher, the injection period, the gas mixture formation period, and the combustion period are effectively shortened. It should be noted that the fuel pressure can be appropriately adjusted according to the properties of the fuel. An upper limit of the fuel pressure may be approximately 120 MPa, for example.

[0176] When the engine speed is low, the time required for the crank angle to change by the same angle is longer than when the engine speed is high. Therefore, shortening the reaction time of the gas mixture through high-pressure delay injection is particularly effective in preventing abnormal combustion. On the other hand, when the engine speed is high, shortening the reaction time of the gas mixture is not particularly effective in preventing abnormal combustion due to the shorter time required for the crank angle to change by the same angle.

[0177] Furthermore, in the high-pressure delay injection, the fuel is injected into the combustion chamber 17 only after it reaches close to CTDC, and a fuel-free gas, in other words, the gas with a high specific heat ratio, is compressed within the combustion chamber 17 during the compression stroke. If the high-pressure delay injection is performed when the engine speed is high, the temperature in the combustion chamber 17 at CTDC, that is, the compression end temperature, rises, which may cause abnormal combustion such as knocking.

[0178] Therefore, in the engine 1, the high-load range (C) is divided into a first high-load segment (C1) on the low engine speed side and a second high-load segment (C2) where the engine speed is higher than within the first high-load segment (C1). If the high-load range (C) is evenly divided into three ranges of low engine speed, medium engine speed, and high engine speed, the first high-load segment (C1) may include the low engine speed range and the medium engine speed range, and the second high-load segment (C2) may include the high engine speed range.

[0179] Within the first high-load segment (C1), the injector 6 performs the above-described high-pressure delay injection in response to receiving the control signal from the ECU 10. Within the second high-load segment (C2), the injector 6 performs fuel injection at a given timing of the intake stroke in response to receiving the control signal from the ECU 10. The fuel injection performed during the intake stroke does not require high fuel pressure. Therefore, the ECU 10 outputs the control signal to the fuel supply system 61 so that the fuel pressure drops below the high-pressure delay injection fuel pressure (e.g., below approximately 40 MPa). Since lowering the fuel pressure reduces mechanical resistance loss of the engine 1, it is beneficial for improving fuel efficiency.

[0180] The specific heat ratio of the gas within the combustion chamber 17 decreases by injecting fuel into the combustion chamber 17 during the intake stroke, thus lowering the compression end temperature and thus preventing abnormal combustion. Since it is not necessary to retard the ignition timing to prevent abnormal combustion, the spark plug 25 ignites the gas mixture at a time close to CTDC within the second high-load segment (C2), similar to the first high-load segment (C1).

[0181] Since the gas mixture does not result in auto-ignition in the first high-load segment (C1) due to the application of the high-pressure delay injection, the engine 1 performs stable SI combustion. Since the gas mixture does not result in auto-ignition in the second high-load segment (C2) because the fuel is injected during the fuel injection stroke, the engine 1 performs stable SI combustion. <spcci-verbrennung>

[0182] Next, the SPCCI combustion described above is described in detail. The upper graph of Fig. 8 illustrates a waveform 801 that is an example of a change in a heat generation rate with respect to the crank angle. When the spark plug 25 ignites the gas mixture near CTDC, specifically at a given time before CTDC, combustion begins by flame propagation. Heat generation in SI combustion is slower than heat generation in CI combustion. Therefore, the waveform of the heat generation rate has a relatively gentle slope. Although not shown, a pressure change (dp / dθ) in the combustion chamber 17 is gentler in SI combustion than in CI combustion.

[0183] As the temperature and pressure within the combustion chamber 17 increase due to the SI combustion, the unburned gas mixture ignites itself. In the example of waveform 801, the slope of the heat generation rate waveform changes from gentle to steep, that is, the heat generation rate waveform has an inflection point at a time when the CI combustion begins.

[0184] After CI combustion begins, SI combustion and CI combustion are performed in parallel. Since the heat generation in CI combustion is greater than in SI combustion, the heat generation rate becomes relatively high. It should be noted that since CI combustion is performed after CTDC, piston 3 is moved downward by the rotation of the engine. This prevents the heat generation rate gradient from CI combustion from becoming excessively steep. The dp / dθ in CI combustion thus becomes relatively smooth.

[0185] The dp / dθ can be used as an index that expresses combustion noise. Since SPCCI combustion is capable of reducing the dp / dθ as described above, it is possible to prevent combustion noise from becoming excessively loud. Thus, combustion noise is reduced below an acceptable level.

[0186] SPCCI combustion ends by terminating CI combustion. CI combustion has a shorter combustion duration than SI combustion. SPCCI combustion advances the combustion end timing compared to SI combustion. In other words, SPCCI combustion brings the combustion end timing closer to CTDC during the expansion stroke. SPCCI combustion is more beneficial than SI combustion in improving the fuel efficiency of engine 1.

[0187] Therefore, SPCCI combustion achieves both prevention of combustion noise and improvement of fuel efficiency. (SI ratio: heat quantity ratio)

[0188] Here, an SI ratio is defined as a parameter that indicates a characteristic of SPCCI combustion. The SI ratio is defined as an index that refers to a ratio of the amount of heat generated by SI combustion with respect to a total amount of heat generated by SPCCI combustion. The SI ratio is a heat ratio resulting from two combustions with different combustion modes. The SI ratio may be a ratio of the amount of heat generated by SI combustion with respect to the amount of heat generated by SPCCI combustion. For example, in waveform 801, the SI ratio may be expressed as SI ratio = (area of ​​SI combustion) / (area of ​​SPCCI combustion). In waveform 801, the SI ratio may be referred to as an "SI fuel ratio" in the sense of the ratio of fuel to be burned during SI combustion.

[0189] In SPCCI combustion, which combines SI combustion and CI combustion, the SI ratio is the ratio of SI combustion to CI combustion. The SI combustion ratio is high when the SI ratio is high, and the CI combustion ratio is high when the SI ratio is low.

[0190] Various definitions can be considered for the SI ratio without limiting the definition described above. For example, the SI ratio can be a ratio of the amount of heat generated by SI combustion with respect to the amount of heat generated by CI combustion. In other words, Fig. 8 the SI ratio = (area of ​​SI combustion) / (area of ​​CI combustion) can be set.

[0191] Furthermore, in SPCCI combustion, the waveform of the heat generation rate has an inflection point at the time when CI combustion starts. Therefore, as indicated by a reference numeral 802 in the middle graph of Fig. 8, by providing a boundary at the inflection point in the heat generation rate waveform, SI combustion may be applied to a region on the advance side of the boundary, and CI combustion may be applied to a region on the retard side of the boundary. In this case, as indicated by hatching the waveform 802, based on an area Q SI of the area of ​​the forward side and an area Q CI of the delay side area the SI ratio = Q SI / (Q SI +Q CI ) or SI ratio =Q SI / Q CI Alternatively, the SI ratio can be defined based on an area of ​​a portion of the advance side region and an area of ​​a portion of the retard side region, instead of the entire area.

[0192] Instead of defining the SI ratio based on heat generation, based on a crank angle Δθ SI the area of ​​the forward side and a crank angle Δθ CI of the delay side area the SI ratio = Δθ SI / (Δθ SI +Δθ CI ) or SI ratio = Δθ SI / Δθ CI be set.

[0193] Based on a peak ΔP SI the heat generation rate in the region of the advance side and a peak ΔP CI the heat generation rate in the area of ​​the delay side, the SI ratio = ΔP SI / (ΔP SI +ΔP CI ) or SI ratio = ΔP SI / ΔP CI be set.

[0194] Furthermore, based on a gradient φ SI the heat generation rate in the region of the advance side and a slope φ CI the heat generation rate in the area of ​​the delay side, the SI ratio = φ SI / (φ SI +φ CI ) or SI ratio =φ SI / φ CI be set.

[0195] Additionally, in this embodiment, the SI ratio is defined by one of area (i.e., the heat generation amount), length along the horizontal axis (i.e., the crank angle), length along the vertical axis (i.e., the heat generation rate), and slope (i.e., the rate of change of the heat generation rate) based on the waveform of the heat generation rate. Although not shown, the SI ratio may be similarly defined based on a waveform of the pressure (P) in the combustion chamber 17, by one of area, length along the horizontal axis, length along the vertical axis, or slope.

[0196] In SPCCI combustion, the inflection point of the combustion waveform with respect to the heat generation rate or pressure does not necessarily always appear clearly. The following can be used as a definition of the SI ratio, which is not based on the inflection point. That is, as indicated by reference numeral 803 in the lower graph of Fig. 8, in the combustion waveform, SI combustion can be applied to a region on the advance side of CTDC, and CI combustion can be applied to a region on the retard side of CTDC. Under this condition, the SI ratio can be determined by a value determined by area (Q SI , Q CI ), length in the horizontal axis (Δθ SI , Δθ CI ), length in the vertical axis (ΔP SI , ΔP CI ) and the slope (φ SI , φ CI ) can be defined.

[0197] Alternatively, the SI ratio may be defined based on the fuel quantity rather than the actual combustion waveform in the combustion chamber 17. As described later, within the mid-load range where SPCCI combustion is performed, split injections including first-stage injection and second-stage injection may be performed. The fuel injected into the combustion chamber 17 by the second-stage injection ignites within a short time after the injection, reaching the vicinity of the spark plug 25 without spreading within the combustion chamber 17. Therefore, the fuel injected into the combustion chamber 17 by the second-stage injection burns mainly in SI combustion.

[0198] On the other hand, the fuel injected into the combustion chamber 17 by the first-stage injection burns mainly during CI combustion. Therefore, the SI ratio can be defined based on the amount of fuel injected during the first-stage injection (m1) and the amount of fuel injected during the second-stage injection (m2). In other words, the SI ratio can be set to m2 / (m1+m2) or m2 / m1. (Stabilization of SPCCI combustion)

[0199] To properly perform SPCCI combustion, the SI combustion must be stabilized. The combustion containing the CI combustion will not stabilize if the SI combustion is unstable.

[0200] A factor related to the stability of SI combustion is the rate of turbulent combustion. When the rate of turbulent combustion is high, SI combustion stabilizes. The rate of turbulent combustion is influenced by the air-fuel ratio (or excess air ratio λ) of the gas mixture, the gas / fuel ratio of the gas mixture, the temperature and pressure in the combustion chamber, the kinetic energy of turbulence in the combustion chamber, etc.

[0201] According to the study of the inventors of the present invention, it was confirmed that the SI combustion stabilizes when λ of the gas mixture is about 1.0 ± about 0.2, the range of G / F of the gas mixture is about 18:1 to about 30:1, the range of the required temperature T IG within the combustion chamber at the time of ignition is about 570 to about 800 K, the range of the required pressure P IG within the combustion chamber at the time of ignition is about 400 to about 920 kPa and the range of kinetic turbulence energy in the combustion chamber is about 17 to about 40 m 2 / m 2 is.

[0202] It was also confirmed that it is possible to extend the range of G / F of the gas mixture to about 50, which is above 30, by stratifying the gas mixture in the combustion chamber.

[0203] That is, SI combustion in SPCCI combustion is achieved by the spark plug 25 igniting the gas mixture. The gas mixture near the spark plug 25 primarily burns in SI combustion. Therefore, the state of the fuel spray moving in the combustion chamber 17 is controlled, for example, by using swirl flow. In this way, even if the G / F range in the combustion chamber at the ignition timing exceeds 30, the G / F of the gas mixture near the spark plug 25 is made relatively lower (stratified) than the gas mixture located away from the spark plug 25, for example, between about 14 and about 22.

[0204] Thus, even if the gas mixture is advantageously diluted to improve fuel efficiency, the SI combustion is stabilized, and the SPCCI combustion is properly performed. <Betriebssteuerung bzw. -regelung des Motors in Lastrichtung>

[0205] By adopting the operating range map 700, the engine 1 switches between SI combustion and SPCCI combustion according to the operating state. Furthermore, the engine 1 changes the SI ratio according to the operating state of the engine 1. Thus, the engine 1 prevents the generation of combustion noise and improves fuel efficiency.

[0206] Fig. Figure 9 is a graph showing a change in the SI ratio, a change in the state function within the combustion chamber 17, changes in the opening periods of the intake valve 21 and the exhaust valve 22, and changes in the fuel injection timing and ignition timing with respect to the engine load. The following describes the operation control of the engine 1 for a condition in which the engine load gradually increases at a given engine speed. (Low load range (low load SI combustion))

[0207] Within the low-load range (A), engine 1 performs low-load SI combustion. When the operating condition of engine 1 is within the low-load range (A), the SI ratio is constant at 100%.

[0208] In the low-load range (A), as described above, the G / F of the gas mixture is fixed between approximately 18:1 and approximately 50:1. The engine 1 introduces the fresh air and the combusted gas into the combustion chamber 17 in amounts corresponding to the amount of fuel. The introduction amount of fresh air, as described above, is changed by throttling and / or the mirror cycle. Since the dilution ratio is high, the temperature within the combustion chamber 17 is increased to stabilize the SI combustion. Within the low-load range (A), the engine 1 introduces the internal EGR gas into the combustion chamber 17.

[0209] The internal EGR gas is introduced into the combustion chamber 17 (i.e., the burned gas is confined within the combustion chamber 17) by providing a negative overlap period in which the intake and exhaust valves 21 and 22 are both closed above the exhaust TDC. Adjustment of the amount of internal EGR gas is performed by appropriately adjusting the length of the negative overlap period, by the electrically actuated intake S-VT 23 changing the opening timing of the intake valve 21, and by the electrically actuated exhaust S-VT 24 changing the opening timing of the exhaust valve 22.

[0210] It should be noted that the internal EGR gas can be introduced into the combustion chamber 17 by providing a positive overlap period in which the intake and exhaust valves 21 and 22 are both opened.

[0211] That is, a period during which the opening period of the intake valve 21 and the opening period of the exhaust valve 22 overlap is provided. In this way, the high-temperature combusted gas in the combustion chamber 17 partially flows into the intake port 18, which communicates with the combustion chamber 17, by opening the intake valve 21. The combusted gas flowing into the intake port 18 is reintroduced into the combustion chamber 17 during the intake stroke. The amount of internal EGR gas can be changed by changing the length of the positive overlap period and the timing of the positive overlap period.

[0212] Note that introducing internal EGR gas by setting the positive overlap period advances the opening period of intake valve 21 compared to introducing internal EGR gas by setting the negative overlap period. Advancing the opening time of intake valve 21 increases the effective compression ratio and raises the temperature in combustion chamber 17. Therefore, it becomes possible to set a low geometric compression ratio of engine 1 while ensuring the stability of SPCCI combustion according to the temperature rise. Thus, cooling loss, mechanical loss, pumping loss, etc., are achieved.

[0213] Furthermore, introducing internal EGR gas by setting a positive overlap period reduces the combustion chamber temperature more than introducing internal EGR gas by setting a negative overlap period, during which the high-temperature combusted gas is confined as it is in the combustion chamber. As a result, auto-ignition during SPCCI combustion is reduced and excessive CI combustion is prevented.

[0214] Within the low-load range (A), the charge amount into the combustion chamber 17 is set to be below 100%. The amount of fresh air introduced into the combustion chamber 17 and the amount of internal EGR gas gradually increase as the fuel quantity increases. For example, the EGR ratio within the low-load range (A) is approximately 40%.

[0215] The injector 6 injects fuel into the combustion chamber 17 during the intake stroke. A homogeneous gas mixture forms within the combustion chamber 17, with an excess air ratio λ of approximately 1.0 ± approximately 0.2 and a G / F ratio of approximately 18:1 to approximately 50:1. The excess air ratio λ is preferably approximately 1.0 to approximately 1.2. By igniting the gas mixture at the given time before CTDC, the gas mixture burns by flame propagation without achieving autoignition. (Second medium load segment (non-boosted SPCCI combustion))

[0216] As the engine load increases and the operating condition enters the second mid-load segment (B2), engine 1 switches from low-load SI combustion to non-boosted SPCCI combustion. The SI ratio drops below 100%. The fuel quantity increases with increasing engine load. When the engine load is low within the second mid-load segment (B2), the CI combustion ratio increases with increasing fuel quantity. The SI ratio gradually decreases with increasing engine load. In the example of Fig. 9, the SI ratio decreases to a given value (lowest value) G1 of less than about 50%.

[0217] As the fuel quantity increases, the combustion temperature rises within the second mid-load segment (B2). If the temperature in the combustion chamber 17 rises excessively, heat generation at the beginning of CI combustion becomes sharp, resulting in an increase in combustion noise.

[0218] Therefore, within the second mid-load segment (B2), the ratio between the internal EGR gas and the external EGR gas is changed according to the change in engine load to adjust the temperature within the combustion chamber 17 before compression starts. That is, as the engine load increases, the internal EGR gas is gradually reduced, and the cooled external EGR gas is gradually increased. Within the second mid-load segment (B2), the negative overlap period changes from a longest length to zero as the engine load increases. Also, within the second mid-load segment (B2), the internal EGR gas becomes approximately zero when the engine load reaches a highest value.

[0219] This is similar when the positive overlap period is provided between the intake and exhaust valves 21 and 22 to introduce the internal EGR gas into the combustion chamber 17. As a result of adjusting the temperature within the combustion chamber 17 by adjusting the overlap period, the SI ratio in SPCCI combustion is adjusted.

[0220] Within the second mid-load segment (B2), the opening of the EGR valve 54 is changed so that the external EGR gas increases as the engine load increases. The amount of external EGR gas introduced into the combustion chamber 17, when expressed by the EGR ratio, is adjusted, for example, between about 0 and about 30%. Within the second mid-load segment (B2), the EGR gas is exchanged from the internal EGR gas to the external EGR gas as the engine load increases. Since the temperature within the combustion chamber 17 is also controlled by adjusting the EGR ratio, the SI ratio of the SPCCI combustion is changed.

[0221] It should be noted that the amount of EGR gas introduced into the combustion chamber 17 is constant between the low-load range (A) and the second medium-load segment (B2). In a low-engine-load section of the second medium-load segment (B2), a large amount of internal EGR gas is introduced into the combustion chamber 17, just like the low-load range (A). Since the temperature in the combustion chamber 17 becomes high, the gas mixture ignites reliably when the engine load is low. In a high-engine-load section of the second medium-load segment (B2), the external EGR gas is introduced into the combustion chamber 17. Since the temperature in the combustion chamber 17 decreases, the combustion noise associated with CI combustion is reduced when the load is high.

[0222] Within the second mid-load segment (B2), the charge amount introduced into the combustion chamber 17 is set to approximately 100%. The throttle valve 43 is essentially fully open. By changing the EGR gas amount, which is the total of the internal EGR gas and the external EGR gas, the amount of fresh air introduced into the combustion chamber 17 is changed to the amount corresponding to the fuel amount.

[0223] As the CI combustion ratio increases during non-boosted SPCCI combustion, the auto-ignition timing advances. As the auto-ignition timing moves forward to CTDC, heat generation at the beginning of CI combustion becomes sharp, resulting in an increase in combustion noise. Therefore, once the engine load reaches the given load G1, the SI ratio for engine 1 gradually increases as the engine load continues to increase.

[0224] That is, engine 1 increases the SI combustion ratio as the fuel quantity increases. For example, as shown in the upper graph of Fig. As shown in Figure 10, in non-boosted SPCCI combustion, the ignition timing is gradually advanced as the fuel amount increases. Since the temperature inside the combustion chamber 17 is adjusted by reducing the introduction amount of the internal EGR gas and increasing the introduction amount of the external EGR gas as described above, the temperature rise at CTDC is suppressed even if the SI ratio is increased with increasing fuel amount. The slope of the heat generation rate of SI combustion hardly changes even if the engine load increases. Advancing the ignition timing causes SI combustion to start earlier, and the heat generation amount of SI combustion increases accordingly.

[0225] As a result of suppressing the temperature rise in the combustion chamber 17 caused by SI combustion, the unburned gas mixture self-ignites at a time point after CTDC. The amount of heat generated by CI combustion is substantially the same even when the engine load increases because the amount of heat generated by SI combustion is increased. Therefore, by gradually setting the SI ratio higher in accordance with the increase in engine load, combustion noise is prevented from increasing. Note that the center of gravity of combustion in non-boosted SPCCI combustion slows down as the engine load increases.

[0226] Within the second mid-load segment (B2), during the compression stroke, the injector 6 injects fuel into the combustion chamber 17 in two stages: first-stage injection and second-stage injection. During the first-stage injection, the fuel is injected at a time separate from the ignition timing, and during the second-stage injection, the fuel is injected at a time close to the ignition timing. (First medium-load segment (boosted SPCCI combustion))

[0227] As the engine load continues to increase and the operating state of the engine 1 enters the first mid-load segment (B1), the booster 44 boosts the fresh air and the external EGR gas. The amount of fresh air introduced into the combustion chamber 17 and the amount of external EGR gas both increase as the engine load increases. The amount of external EGR gas introduced into the combustion chamber 17, when expressed by the EGR ratio, is, for example, about 30%. The EGR ratio is substantially constant regardless of the engine load. Therefore, the G / F of the gas mixture is also substantially constant regardless of the engine load. Note that the amount of EGR gas introduced into the combustion chamber 17 is continuous between the second mid-load segment (B2) and the first mid-load segment (B1).

[0228] The SI ratio is fixed at a given value below approximately 100% with respect to the engine load. When comparing the SI ratio of the second medium-load segment (B2), specifically the SI ratio that gradually increases as the engine load increases from a value above the given load G1, with the SI ratio of the first medium-load segment (B1), the SI ratio of the first medium-load segment (B1), where the engine load is higher, is higher than that of the second medium-load segment (B2). The SI ratio is continuous across the boundary between the first medium-load segment (B1) and the second medium-load segment (B2).

[0229] The SI ratio within the first medium-load segment (B1) may be slightly changed according to the change in the engine load. The rate of change of the SI ratio according to the engine load within the first medium-load segment (B1) may be lower than that on the high engine load side of the second medium-load segment (B2).

[0230] As shown in the graphic below from Fig. As shown in Figure 10, even in boosted SPCCI combustion, the ignition timing is gradually advanced as the fuel quantity increases. Since the amount of fresh air and EGR gas introduced into the combustion chamber 17 is increased by boosting as described above, the amount of heat is large. Therefore, even if the fuel quantity increases, the temperature rise in the combustion chamber 17 caused by SI combustion is suppressed. The waveform of the heat generation rate of boosted SPCCI combustion becomes larger (the area of ​​the section formed by the waveform and the horizontal axis become larger) in a similar shape as the engine load increases.

[0231] That is, the heat generation amount of SI combustion increases, while the slope of the heat generation rate of SI combustion hardly changes. The unburned gas mixture self-ignites at essentially the same time after CTDC. The heat generation amount of CI combustion increases as the engine load increases. As a result, within the first mid-load segment (B1), since both the heat generation amount of SI combustion and the heat generation amount of CI combustion increase, the SI ratio is constant with respect to the engine load. Although the combustion noise increases as the peak heat generation of CI combustion increases, a certain level of combustion noise is permitted because the engine load within the first mid-load segment (B1) is relatively high. Note that the center of combustion in boosted SPCCI combustion slows down as the engine load increases.

[0232] Within the first mid-load segment (B1), the overlap period in which the intake and exhaust valves 21 and 22 are both open (positive overlap period) is provided above the exhaust TDC. The unburned gas in the combustion chamber 17 is purged by the boost pressure. Thus, the temperature in the combustion chamber 17 drops, and as a result, abnormal combustion is prevented from occurring when the engine load is relatively high. Further, by lowering the temperature in the combustion chamber 17 within the range where the engine load is relatively high, the auto-ignition timing is adjusted to an appropriate timing and the SI ratio is maintained at a given SI ratio. Furthermore, by purging the burned gas, the charge amount of fresh air in the combustion chamber 17 is increased.

[0233] Within the first mid-load segment (B1), the injector 6 injects fuel into the combustion chamber 17 in two stages, similar to the second mid-load segment B2): first-stage injection and second-stage injection, and during the compression stroke. During the first-stage injection, the fuel is injected at a time separate from the ignition timing, and during the second-stage injection, the fuel is injected at a time close to the ignition timing. The first-stage injection can be performed in the period from the intake stroke to the early half of the compression stroke, and the second-stage injection can be performed in the period from the latter half of the compression stroke to the early half of the expansion stroke.

[0234] By performing the first-stage injection and the second-stage injection during these periods, the injector 6 forms a substantially homogeneous gas mixture in the combustion chamber 17, with the excess air ratio λ being approximately 1.0 ± approximately 0.2 and the G / F being approximately 18:1 to approximately 30:1. Since the gas mixture is substantially homogeneous, fuel efficiency can be improved by reducing unburned fuel loss and exhaust performance can be improved by preventing smoke generation. The excess air ratio λ is preferably approximately 1.0 to approximately 1.2.

[0235] By igniting the gas mixture at the given timing before the CTDC, the gas mixture burns by flame propagation. After this combustion begins, the unburned gas mixture ignites itself and causes CI combustion. The fuel injected during the second-stage injection mainly causes SI combustion. The fuel injected during the first-stage injection mainly causes CI combustion. Since the first-stage injection is performed on the compression stroke, the fuel injected during the first-stage injection is prevented from causing abnormal combustion such as pre-ignition. Furthermore, the fuel injected during the second-stage injection is stably burned by flame propagation. (High-load range (high-load SI combustion))

[0236] As the engine load continues to increase and the operating condition of engine 1 enters the high-load range (C), engine 1 performs high-load SI combustion. Therefore, the SI ratio within the high-load range (C) becomes approximately 100%.

[0237] The throttle valve 43 is substantially fully open. The booster 44 boosts the fresh air and the external EGR gas even within the high-load range (C). The EGR valve 54 gradually decreases the introduction amount of the external EGR gas by adjusting its opening as the engine load increases. In this way, the fresh air introduced into the combustion chamber 17 increases as the engine load increases. Since the fuel amount increases as the fresh air amount increases, it is advantageous to increase the peak power of the engine 1. It should be noted that the amount of EGR gas introduced into the combustion chamber 17 is continuous between the first medium-load segment (B1) and the high-load range (C).

[0238] Also within the high-load range (C), similar to the first medium-load segment (B1), the overlap period, in which the intake and exhaust valves 21 and 22 are both open, is provided above the exhaust TDC. The unburned gas in the combustion chamber 17 is purged by the boost pressure. This prevents the occurrence of abnormal combustion. Furthermore, the amount of fresh air charged into the combustion chamber 17 is increased.

[0239] In a low engine speed segment of the high-load range (C) (i.e., the first high-load segment (C1)), the injector 6 injects fuel into the combustion chamber 17 in the delay period as described above. In a high engine speed segment of the high-load range (C) (i.e., the second high-load segment (C2)), the injector 6 injects fuel into the combustion chamber 17 during the intake stroke. Within each segment, a substantially homogeneous gas mixture, in which the excess air ratio λ is about 1.0 ± about 0.2 and the G / F is about 18:1 to about 50:1, is formed within the combustion chamber 17.

[0240] At the highest load, the excess air ratio λ is, for example, 0.8. Furthermore, the G / F of the gas mixture can be approximately 17:1, for example, at the highest load. By igniting the gas mixture at the given time before the CTDC, the spark plug 25 burns the gas mixture by flame propagation. Within the high-load range (C), the gas mixture causes SI combustion without achieving autoignition due to high-pressure delay injection or fuel injection during the intake stroke. <Betriebssteuerung bzw. -regelung des Motors in Drehzahlrichtung> (Adjusting the SI ratio)

[0241] Fig. Figure 11 shows a relationship between the engine speed and the SI ratio at a given load within the medium load range (B) in which the SPCCI combustion is performed. As shown in the operating range map 700 of Fig. As shown in Figure 7A, the speed N1 is a lowest speed within the medium-load range (B). A speed N2 is a highest speed within the medium-load range (B), which lies at a boundary to the high-speed range (D) in which the SI combustion is performed (this is the case in the following description).

[0242] CI combustion is disadvantageous in that the combustion noise is louder than SI combustion. However, when the engine speed is low, the NVH of engine 1 is generally low. Therefore, at low engine speed, even by adding the combustion noise of CI combustion to the NVH, NVH is kept lower than the allowable value. Therefore, at low engine speed, the ECU 10 lowers the SI ratio, and CI combustion is performed sufficiently. Thus, at low engine speed, fuel efficiency is improved without NVH becoming a problem.

[0243] On the other hand, as the engine speed increases, the NVH of engine 1 also increases accordingly. Furthermore, when the combustion noise of the CI combustion is added to the increased NVH, the NVH may exceed the allowable value. Therefore, as the engine speed increases, the ECU 10 increases the SI ratio according to the speed increase.

[0244] As in Fig. As shown in Figure 11, the ECU 10 of this configuration example increases the SI ratio linearly as the engine speed increases. Since the highest speed N2 is the boundary of the high-speed range (D) where only SI combustion is performed, the SI ratio is set to approximately 100% at the N2 speed. Therefore, the transition between the medium-load range (B) where SPCCI combustion is performed and the high-speed range (D) where SI combustion is performed is performed smoothly without a significant change in the combustion state.

[0245] In other words, the CI combustion ratio is continuously reduced by adjusting the SI ratio according to the engine speed, and the engine load range in which self-ignition combustion is performed is expanded as high as possible. As a result, in the engine 1, the operating range in which self-ignition combustion is performed can be expanded not only toward the higher engine load side but also toward the higher engine speed side, while NVH, which can become a problem, is kept below the allowable value. Therefore, the engine 1 excels in fuel efficiency. Note that such adjustment of the SI ratio is achievable by sophisticated control of the engine 1 performed by the ECU 10 (described in detail later).

[0246] When the engine load is high, the amount of fuel increases and the amount of heat absorbed by the combustion chamber 17 also increases, and therefore the temperature of the combustion chamber 17 becomes relatively high. Thus, the SI combustion in SPCCI combustion becomes sharper at high engine load than at low engine load. Therefore, when compared at the same engine speed, the SI ratio is adjusted to be higher at high engine load than at low engine load. As shown by the dot-dash line in Fig. As indicated in Figure 11, the straight line indicating the change in the SI ratio with respect to the engine speed (SI ratio / engine speed) has a gentler slope as the engine load increases. (Method for increasing the SI ratio at high engine speed)

[0247] At high engine speed, the combustion cycle is shorter than at low engine speed. Since the amount of heat absorbed by the inner wall of the combustion chamber, etc., during combustion decreases accordingly, the temperature of the combustion chamber 17 drops compared to when the engine speed is low. The gas mixture also becomes less likely to absorb heat from the combustion chamber 17. In addition, the combustion time of the gas mixture becomes short. At high engine speed, the SI ratio must be increased under such an adverse condition.

[0248] In this regard, by increasing the temperature of the gas mixture itself in the combustion chamber 17, combustion is stimulated even in the state where heat is not easily transferred from the combustion chamber 17.

[0249] Furthermore, the ignition of CI combustion occurs after SI combustion. Although increasing the gas mixture temperature stimulates both SI combustion and CI combustion, the combustion duration is short when the engine speed is high. Thus, SI combustion becomes sharper and is stimulated more than CI combustion, which starts later than SI combustion. Therefore, increasing the gas mixture temperature increases the SI ratio when the engine speed is high. (First procedure)

[0250] Therefore, in order to change the SI ratio according to the change in engine speed, in the engine 1, the internal EGR ratio is set within the middle load range (B) in which the SPCCI combustion is performed.

[0251] For example, the ECU 10 outputs a control signal to the electrically operated intake S-VT 23 and / or the electrically operated exhaust S-VT 24 to change the opening timing of the intake valve 21 and / or the exhaust valve 22. In this way, the negative overlap period is provided, the length of the period is changed, and the high-temperature burned gas (internal EGR gas) introduced into the combustion chamber 17 is adjusted to a given amount. Thus, an internal EGR ratio, specifically, "a mass ratio (%) of the internal EGR gas with respect to the total gas within the combustion chamber 17," is changed.

[0252] As in Fig. As shown in Figure 7A, in the engine 1 adopting the operating range map 700, the mid-load range (B) is divided into the first mid-load segment (B1) on the higher engine load side and the second mid-load segment (B2) on the lower engine load side. In the first mid-load segment (B1), while corresponding to the high engine load, since the operating range in which SPCCI combustion is performed is expanded to the higher engine load side, boosting is performed, whereas in the second mid-load segment (B2), boosting is not performed.

[0253] In this engine 1, within the first mid-load segment (B1) where boosting is performed, even if the internal EGR gas is introduced, scavenging is performed by the boost pressure because of the positive overlap period. Therefore, within the first mid-load segment (B1), the first method is not applied because it is difficult to change the internal EGR ratio. Within the second mid-load segment (B2), the ECU 10 outputs the control signal to the electrically driven intake S-VT 23, etc., so that the internal EGR ratio becomes higher at a high engine load than at a low engine load.

[0254] Fig. Figure 12A shows a relationship between the engine speed and the internal EGR ratio (internal EGR ratio / engine speed) at given loads within the second mid-load segment (B2). The ECU 10 linearly increases the internal EGR ratio as the engine speed increases from the speed N1, which is the lowest speed. As a result, the temperature of the gas mixture in the combustion chamber 17 rises, and therefore, even if the amount of heat absorbed by the combustion chamber 17 is small and the combustion duration is short, the SI combustion becomes sharp and the SI ratio increases. Since the CI combustion ratio decreases as the SI ratio increases, the combustion noise caused by the SPCCI combustion is reduced. When the engine speed is high, the NVH is suppressed below the allowable value. Note that the relationship between the engine speed and the internal EGR ratio shown in Fig. 12A and Fig. 12B are examples and the rate of change can be suitably changed according to the specifications.

[0255] As in Fig. As shown in Figure 12A, when the engine speed exceeds a given speed (limit start speed Na), the increase in the internal EGR ratio is limited. For example, to change the internal EGR ratio according to the change in engine speed, the memory 102 of the ECU 10 stores a given map, and the limit start speed Na is set in the map. When the engine speed exceeds the limit start speed Na, the ECU 10 outputs the control signal to the electrically operated intake S-VT 23, etc., to limit the increase in the internal EGR ratio.

[0256] Note that the limit start speed Na is set at least higher than a middle value of the engine speed range (between the speed N1 and the speed N2) within the medium load range (B). Particularly, when the engine speed range within the medium load range (B) is evenly divided into three segments, the speed Na is preferably set within the high speed segment.

[0257] In this configuration example, the internal EGR ratio peaks once it exceeds the limit start speed Na and remains essentially constant (depending on specifications, the value may vary slightly). As the internal EGR ratio increases and the gas mixture temperature rises excessively, CI combustion also becomes sharp and combustion noise increases, which may cause NVH exceeding the allowable level. Therefore, the ECU 10 sets the limit start speed Na lower than the value that may cause such a situation to suppress NVH below the allowable level.

[0258] When the engine load is high, the fuel quantity increases and the heat absorption amount of the combustion chamber 17 also increases, and therefore the temperature of the combustion chamber 17 rises relatively. Thus, the temperature of the gas mixture becomes higher, and the SI combustion in SPCCI combustion becomes sharper at a high engine load than at a low engine load. For this reason, the ECU 10 changes the internal EGR ratio across the entire engine speed range as the engine load increases.

[0259] Furthermore, the limit start speed Na is set lower as the engine load increases. Since the gas mixture temperature tends to increase excessively as the engine load increases, setting the limit start speed Na lower as the engine load increases more reliably prevents NVH from exceeding the allowable value.

[0260] Fig. 12A shows a relationship between the engine speed and the internal EGR ratio in the case of introducing the internal EGR gas by setting the negative overlap period (also referred to as "NVO setting"). Fig. 12B shows a relationship between the engine speed and the internal EGR ratio in the case of introducing the internal EGR gas by setting the positive overlap period (also referred to as "PVO setting").

[0261] When the internal EGR gas is introduced through the PVO setting, unlike the internal EGR gas being introduced through the NVO setting, which keeps the high-temperature burned gas as it is in the combustion chamber 17, the high-temperature burned gas flows out once toward the intake port 18 and is again introduced into the combustion chamber 17. The high-temperature burned gas introduced into the combustion chamber 17, that is, the internal EGR gas, is cooled by flowing toward the intake port 18 where the temperature is low.

[0262] Therefore, when the internal EGR gas is introduced through the PVO setting, the temperature of the combustion chamber 17 becomes relatively lower compared to when the internal EGR gas is introduced through the NVO setting. Thus, even if the internal EGR ratio increases, the temperature of the gas mixture is not likely to rise excessively. As a result, the setting of the limit start speed Na becomes unnecessary, or even if it is needed, it is set sufficiently higher than when the NVO is set.

[0263] Therefore, when the internal EGR gas is introduced through the PVO setting, the internal EGR ratio increases steadily without any limitation until the highest speed N2 is reached. Although the limitation start speed Na is close to the highest speed N2 at a high engine load, which Fig. 12B, the setting of the limit start speed Na may be unnecessary even at a high engine load. There may also be a case where the setting of the limit start speed Na is necessary at a low engine load or a medium engine load. Note that even in this case, the limit start speed Na is set closer to the highest speed N2 compared to the introduction of the internal EGR gas by the NVO setting.

[0264] It should be noted that in the engine 1 using an operating range map 701 and an operating range map 702, which will be described later, a swirl flow with a higher intensity than a given intensity is formed in the combustion chamber 17. With such a swirl flow formed in the combustion chamber 17, the gas mixture in the combustion chamber 17 is stratified by the flow. Thus, the gas mixture near the spark plug 25 is made suitable for SI combustion, and as a result, stable SPCCI combustion is achieved.

[0265] Therefore, in this case, even if the internal EGR ratio is increased at a high engine speed to ensure the high SI ratio, the EGR gas is reduced by stratifying the gas mixture near the spark plug 25. Thus, the stability of the ignition in the SI combustion is ensured, and the stable SPCCI combustion is achieved. (Second procedure)

[0266] By the first method alone, in the medium load range (B) where SPCCI combustion is performed, the operating range in which self-ignition combustion is extended to the higher engine speed side becomes only for the second medium load segment (B2). Therefore, in this engine 1, the second method is also applied within the first medium load segment (B1), similarly to the second medium load segment (B2), so that the temperature of the gas mixture itself in the combustion chamber 17 is changed to a higher level as the engine load increases.

[0267] That is, this engine 1 changes the external EGR ratio within the medium load range (B) in which the SPCCI combustion is performed to change the SI ratio according to the change in engine speed.

[0268] For example, the introduction amount of the cooled burned gas (external EGR gas) at a lower temperature than the internal EGR gas is changed by changing the opening of the EGR valve 54, and thus the external EGR ratio, specifically, a "mass ratio (%) of the external EGR gas with respect to the total gas within the combustion chamber 17" is changed.

[0269] The ECU 10 outputs the control signal to the EGR valve 54 so that the external EGR ratio decreases at a lower rate at low engine speed than at high engine speed. The external EGR gas has a lower temperature than the mixed gas in the combustion chamber 17. When the introduction amount of the external EGR gas is large, the temperature of the mixed gas drops accordingly. Therefore, by decreasing the external EGR ratio, the temperature of the mixed gas itself in the combustion chamber 17 is changed to a higher level.

[0270] As in Fig. As shown in Figure 7A, the external EGR gas is introduced within both the first mid-load segment (B1) and the second mid-load segment (B2), and the amount of external EGR gas is particularly large within the first mid-load segment (B1). Therefore, the second method is effective within the first mid-load segment (B1), and by combining the first method, the SI ratio is effectively set to high according to the change in engine speed over the entire mid-load range (B).

[0271] Fig. Figure 13A shows a relationship between engine speed and the external EGR ratio (external EGR ratio / engine speed) at given loads within the mid-load range (B). The solid line indicates the relationship within the first mid-load segment (B1), and the dot-dash line indicates the relationship within the second mid-load segment (B2).

[0272] As described above, in the engine 1 adopting the operating range map 700, the SCV 56 is substantially fully opened in each operating range including the middle load range (B) in which the SPCCI combustion is performed. Fig. Fig. 13A shows such a relationship as described above for the case where substantially no swirl flow is generated in the combustion chamber 17 or the flow is weak even if it is generated.

[0273] The ECU 10 essentially maintains the external EGR ratio until the engine speed reaches a predetermined speed (decrease start speed Nb) from the lowest speed N1. In this way, a high EGR ratio (external EGR ratio) is set at a low engine speed, and therefore, even if the G / F value of the gas mixture to be secured is high, it can be set within a suitable range. Because the dilution ratio of the gas mixture is maintained high, fuel efficiency improves.

[0274] Once the reduction start speed Nb is exceeded, the ECU 10 reduces the external EGR ratio according to the engine load. For example, to change the external EGR ratio according to the change in engine speed, the memory 102 of the ECU 10 stores a given map, and the reduction start speed Nb is set in advance in the map.

[0275] The reduction start speed Nb, like the limit start speed Na, is set at least higher than a middle value of the engine speed range within the medium load range (B). Particularly, when the engine speed range within the medium load range (B) is evenly divided into three segments, the reduction start speed Nb is preferably set equal to the limit start speed Na.

[0276] Based on the map, the ECU 10 outputs the control signal to the EGR valve 54 so that the external EGR ratio starts to decrease depending on the engine load once it exceeds the reduction start speed Nb. Note that when the engine load is low, the value of the external EGR ratio can be maintained as it is even if the reduction start speed Nb is exceeded.

[0277] Thus, in a higher speed range at or above the reduction start speed Nb, the ratio of the external EGR gas introduced into the combustion chamber 17 decreases, and the temperature of the gas mixture itself increases. Therefore, even if the heat absorption amount from the combustion chamber 17 is small and the combustion duration is shortened, the SI combustion becomes sharp and the SI ratio increases. Since the CI combustion decreases with increasing SI ratio, the combustion noise generated by the SPCCI combustion is reduced. When the engine speed is high, the NVH is suppressed below the allowable value. It should be noted that the relationship between the engine speed and the external EGR ratio described in Fig. 13A and Fig. 13B is an example, and the change rate thereof can be appropriately changed according to specifications.

[0278] The ECU 10 outputs the control signal to the EGR valve 54 so that the external EGR ratio decreases more with increasing engine load (the reduction rate increases).

[0279] As engine load increases, the fuel quantity increases accordingly. If the amount of burned gas is excessively large relative to the increasing fuel, spark ignition combustion, i.e., the SI combustion in SPCCI combustion, becomes unstable, and the SI ratio may decrease. Therefore, since the rate of reduction of the external EGR ratio is increased as engine load increases to stabilize the ignition, a high SI ratio is maintained.

[0280] The ECU 10 also changes the external EGR ratio within the second mid-load segment (B2) similarly to the first mid-load segment (B1). In this configuration example, the EGR gas introduced within the first mid-load segment (B1) contains only the external EGR gas, while the internal EGR gas is also introduced within the second mid-load segment (B2). Therefore, within the second mid-load segment (B2), the external EGR ratio is changed by considering the state function within the combustion chamber 17, such as the G / F of the gas mixture, due to the relationship with the introduction amount of the internal EGR gas. As a result, the external EGR ratio within the second mid-load segment (B2) decreases relative to that within the first mid-load segment (B1).

[0281] In the engine 1 using one of the operating range map 701 and the operating range map 702 described later, unlike the engine 1 using the operating range map 700, a swirl flow stronger than a given strength is formed in the combustion chamber 17. Fig. 13B illustrates the relationship described above for such an engine 1.

[0282] By forming a swirl flow stronger than the specified strength within the combustion chamber 17, the fluidity of the gas mixture in the combustion chamber 17 is improved. Thus, even when the amount of EGR gas is large, the amount of EGR gas around the spark plug 25 is reduced when ignition is performed. Therefore, even when the engine load increases and the fuel quantity increases, the stability of the ignition combustion is ensured, and therefore, the SI combustion in the SPCCI combustion is performed stably.

[0283] As in Fig. Thus, as shown in FIG. 13B, the ECU 10 outputs the control signal to the EGR valve 54 so that the rate of reduction of the external EGR ratio is kept substantially constant with respect to the change in the engine load. Even when the engine load increases, the ECU 10 reduces the external EGR ratio in the same state as when the engine load is low. (Third procedure)

[0284] In the method of increasing the gas mixture temperature as in the first and second methods, if the temperature rises excessively, the CI combustion becomes harsh and the combustion noise increases, which may cause the NVH to exceed the allowable value. Since the internal EGR ratio and the external EGR ratio are limited by other state functions, such as the gas mixture G / F, simply changing the internal EGR ratio and the external EGR ratio limits the effect of reducing combustion noise in SPCCI combustion, and the effect may not be sufficiently maintained.

[0285] Specifically, in the first method, the increase of the internal EGR ratio is limited within the range above the limit start speed Na. Even in the second method, if the external EGR ratio is excessively reduced within the range above the reduction start speed Nb, the temperature of the gas mixture becomes excessively high and the combustion noise increases, which may cause NVH to exceed the allowable value, and therefore the reduction rate is limited. Therefore, the SI ratio cannot increase to a desired SI ratio within the ranges above the limit start speed Na and the reduction start speed Nb. In this case, the effect of reducing combustion noise is limited, and a sufficient effect cannot be obtained.

[0286] Therefore, in this engine 1, a method other than the method for increasing the temperature of the gas mixture is used, so that the SI ratio is also increased thereby.

[0287] That is, in conjunction with the first or second method, the ECU 10 uses a third method in which the SCV 56 is controlled to increase the intake air flow. As the intake air flow is increased, the SI ratio increases because the SI combustion becomes sharp. As a result, the combustion noise of the SPCCI combustion is reduced.

[0288] Fig. 14A shows a relationship between the rotational speed of the engine 1 using the operating range map 700 and the SCV 56. As described above, in the engine 1 using the operating range map 700, the SCV 56 is substantially fully open in each operating range including the middle load range (B) in which the SPCCI combustion is performed. Fig. Figure 14A illustrates the above relationship in such a case. On the other hand, Fig. 14B the above relationship of the engine 1 having an operating range map 701 of Fig. 7B, and engine 1, which has an operating range map 702 of Fig. 7C is used.

[0289] As in Fig. 14A, in the engine 1 using the operating range map 700, when the engine speed reaches a given speed (operation start speed N4) at which adjustment to the desired SI ratio becomes difficult by the first or second method alone, e.g., the restriction start speed Na at which the restriction of the internal EGR ratio is started, the reduction start speed Nb at which the reduction of the external EGR ratio is started, the ECU 10 outputs the control signal to the SCV 56 to restrict its opening from the substantially fully open state. Thus, the swirl flow in the combustion chamber 17 becomes strong. The ECU 10 linearly reduces the opening of the SCV 56 as the engine speed increases. As the engine speed increases, the SI combustion becomes even sharper because the swirl flow becomes stronger.

[0290] As a result, in the first or second method, even within the higher engine speed range where the limitation is applied, the combustion noise of the SPCCI combustion is reduced. Consequently, when the engine speed is high, NVH is suppressed below the permissible level. By combining the first or second method with the third method, the SI ratio is changed to a higher and more appropriate level.

[0291] On the other hand, as in Fig. 14B, in the engine 1 using the operating range map 701 and the engine 1 using the operating range map 702 described later, the swirl flow stronger than the given strength is formed in the combustion chamber 17 within the region where the SPCCI combustion is performed.

[0292] For example, at the time of reaching the ignition b the intake stroke, the SCV 56 is controlled at a given narrow opening without being fully closed, so that the swirl ratio becomes a range between about 1.5 to about 3. The opening of the SCV 56 or the swirl ratio is appropriately set according to the specifications of the engine 1. Fig. Figure 14B illustrates the case where the SCV 56 is held substantially constant at about 30% with respect to the engine speed.

[0293] When SPCCI combustion is performed, by forming such a swirl flow in the combustion chamber 17, the fluidity in the combustion chamber 17 is improved, and a stratified gas mixture is formed in the combustion chamber 17. Thus, stable SPCCI combustion is achieved within a wide operating range of the engine 1. Details will be described later. (Fourth procedure)

[0294] In the engine 1 adopting the operating range map 700, the SI ratio is increased to the higher speed side also by changing the ignition timing.

[0295] Fig. 15 shows a relationship between the engine speed and the ignition timing at a given load within the middle load range (B) in which SPCCI combustion is performed. Similar to the third method, when the engine speed reaches a predetermined speed (advance start speed Nd) at which adjustment to the desired SI ratio becomes difficult by the first or second method alone (such as the limitation start speed Na at which limitation of the internal EGR ratio is started, or the reduction start speed Nb at which reduction of the external EGR ratio is started), the ECU 10 outputs the control signal to the spark plug 25 to advance the ignition timing. Thus, the start of SI combustion is advanced, and the SI combustion becomes sharp. The ECU 10 advances the ignition timing linearly with increasing engine speed. The SI combustion becomes even sharper as the engine speed increases.

[0296] The fourth method can replace the third method. The fourth method can be used in conjunction with the third method. By appropriately combining the first to fourth methods, adaptation to a stable SI ratio becomes possible even at high engine speeds. (Change in SI ratio)

[0297] Fig. 16 shows a process related to the operation control of the engine 1, which is executed by the ECU 10. Based on the detection signals from the sensors SW1 to SW16, the ECU 10 determines the operating state of the engine 1 and also changes the state function inside the combustion chamber 17, the injection amount, the injection timing, and the ignition timing, so that the combustion in the combustion chamber 17 is of the SI ratio according to the operating state.

[0298] The ECU 10 further adjusts the SI ratio when it determines that its adjustment is needed based on the detection signals from the sensors.

[0299] First, at S1, the ECU reads the detection signals of the sensors SW1 to SW16. Next, at S2, the ECU 10 determines the operating state of the engine 1 based on the detection signals and sets the target SI ratio as shown in Fig. 9 or Fig. 11 is shown.

[0300] Next, at S3, the ECU 10 sets a target in-cylinder state function to achieve the specified target SI ratio based on a preset combustion model. For example, the target temperature, target pressure, and target state function in the combustion chamber 17 are set. At S4, the ECU 10 sets the opening of the EGR valve 54, the opening of the throttle valve 43, the opening of the air bypass valve 48, the phase angles of the electrically operated intake S-VT 23 and the electrically operated exhaust S-VT 24, and the opening of the SCV 56 required to achieve the target in-cylinder state function.

[0301] The ECU 10 sets the control amounts of these devices in advance based on a map stored in the ECU 10. Based on the set control amounts, the ECU 10 outputs control signals to the EGR valve 54, the throttle valve 43, the air bypass valve 48, the electrically operated intake S-VT 23, the electrically operated exhaust S-VT 24, and the SCV 56. When each device operates based on the control signal from the ECU 10, the state function in the combustion chamber 17 becomes the target state function.

[0302] The ECU 10 further calculates a predicted value and an estimated value of the state function in the combustion chamber 17 based on the set control amount of each device. The state function prediction value is obtained by predicting the state function in the combustion chamber 17 before the intake valve 21 closes and is used to adjust the fuel injection amount in the intake stroke, as described later. The state function estimate value is obtained by estimating the state function in the combustion chamber 17 after the intake valve 21 closes and is used to adjust the fuel injection amount and ignition timing in the compression stroke, as described later. The state function estimate value is also used to calculate a state function error based on a comparison with an actual combustion state, as described later.

[0303] At S5, the ECU 10 adjusts the fuel injection amount during the intake stroke based on the predicted state function value. Note that in a case where fuel injection is not performed during the intake stroke, the fuel injection amount is zero. At S6, the ECU 10 controls the injection of the injector 6. That is, the ECU 10 outputs a control signal to the injector 6 to inject fuel into the combustion chamber 17 at a given injection timing.

[0304] At S7, the ECU 10 sets the fuel injection amount for the compression stroke based on the state function estimate and the fuel injection result on the intake stroke. Note that when fuel injection is not performed on the compression stroke, the fuel injection amount is zero. When split injections are performed on the compression stroke, the first-stage injection amount and the second-stage injection amount are respectively adjusted. At S8, the ECU 10 outputs a control signal to the injector 6 to inject fuel into the combustion chamber 17 at the injection timing based on the preset map.

[0305] At S9, the ECU 10 adjusts the ignition timing based on the state function estimate and the fuel injection result during the compression stroke. At S10, the ECU 10 outputs a control signal to the spark plug 25 to ignite the gas mixture in the combustion chamber 17 at the specified ignition timing.

[0306] SI combustion or SPCCI combustion is performed in the combustion chamber 17 by the spark plug 25 igniting the gas mixture. At S11, the ECU 10 reads a pressure change in the combustion chamber 17 detected by the pressure sensor SW6 and determines the combustion state of the gas mixture in the combustion chamber 17 based on this change. At S12, the ECU 10 further compares the combustion state detection result with the state function estimate estimated at S4 to calculate a difference (an error) between the state function estimate and the actual state function. The calculated error is used in the estimation at S4 in the next cycle and / or its subsequent cycles.

[0307] The ECU 10 changes the openings of the throttle valve 43, the EGR valve 54, the SCV 56, and / or the air bypass valve 48, and the phase angles of the electrically operated intake S-VT 23 and the electrically operated exhaust S-VT 24. Thus, the introduction amounts of fresh air and EGR gas into the combustion chamber 17 are changed. The state function error feedback is equivalent to changing the SI ratio when the ECU 10 determines that the SI ratio change is necessary based on the error between the target SI ratio and the actual SI ratio (error).

[0308] If the temperature in the combustion chamber 17 is estimated to be lower than the target temperature based on the state function estimation value at S8, the ECU 10 advances the injection timing on the compression stroke more than the injection timing based on the map, so that the ignition timing advances. On the other hand, if the temperature in the combustion chamber 17 is estimated to be higher than the target temperature based on the state function estimation value at S8, the ECU 10 retards the injection timing on the compression stroke more than the injection timing based on the map, so that the ignition timing retards.

[0309] Therefore, when the temperature in the combustion chamber 17 is low, as indicated by P2 of Fig. 17, the time θ CI In this case, the unburned gas mixture ignites spontaneously after SI combustion, triggered by spark ignition delays, and the SI ratio shifts from the target SI ratio (see P1). In this case, the unburned fuel increases and exhaust performance deteriorates.

[0310] Therefore, if the temperature in the combustion chamber 17 is estimated to be lower than the target temperature, the ECU 10 advances the injection timing and the ignition timing θ IG at S10 from Fig. 16. As stated by P3 in Fig. 17, since the start of the SI combustion is advanced, sufficient heat generation occurs by the SI combustion and therefore the time θ CI The self-ignition of the unburned gas mixture slows down when the temperature in the combustion chamber 17 is low. As a result, the SI ratio approaches the target SI ratio. The increase in unburned fuel and the deterioration of exhaust performance are prevented.

[0311] When the temperature in the combustion chamber 17 is high, as indicated by P4 in Fig. 17, the unburned gas mixture ignites immediately after the start of SI combustion by spark ignition, and the SI ratio shifts from the target SI ratio (see P1). In this case, the combustion noise increases.

[0312] Therefore, if the temperature in the combustion chamber 17 is estimated to exceed the target temperature, the ECU 10 retards the injection timing and the ignition timing θ IG at S10 from Fig. 16. As stated by P5 in Fig. 17, since the start of the SI combustion is delayed; this prevents the time θ C1 The auto-ignition of the unburned gas mixture advances when the temperature in the combustion chamber 17 is high. As a result, the SI ratio approaches the target SI ratio. An increase in combustion noise is avoided.

[0313] The changes in injection timing and ignition timing are equivalent to adjusting the SI ratio when the ECU 10 determines that the SI ratio needs to be adjusted in SPCCI combustion. By adjusting the injection timing, a suitable gas mixture is formed in the combustion chamber 17 at the advanced or retarded ignition timing. The spark plug 25 reliably ignites the gas mixture, and the unburned gas mixture self-ignites at an appropriate time.

[0314] Note that the adjustment of the state function within the combustion chamber 17 is performed by controlling the throttle valve 43, the EGR valve 54, the air bypass valve 48, the electrically operated intake S-VT 23, the electrically operated S-VT exhaust valve 24, and the SCV 56 based on the actual combustion state of. Fig. 17 as in S12 and S4 of Fig. 16 described.

[0315] The engine 1 changes the SI ratio through the state function adjustment devices including the throttle valve 43, the EGR valve 54, the air bypass valve 48, the electrically operated intake S-VT 23, the electrically operated exhaust S-VT 24, and the SCV 56. The SI ratio can be roughly changed by changing the state function within the combustion chamber 17. At the same time, the engine 1 adjusts the SI ratio by adjusting the fuel injection timing and ignition timing. By adjusting the injection timing and ignition timing, for example, the difference between the cylinders can be corrected and the auto-ignition timing is finely adjusted. By adjusting the SI ratio in two stages, the engine 1 accurately achieves the target SPCCI combustion according to the operating condition. (Second operating range map)

[0316] Fig. 7B shows a second configuration example of the operating range map of engine 1 (second operating range map 701). The operating range map 701 is divided into the following three areas. (A): a low-load range including idle operation (B): a medium load range between the low load range (A) and the following high load range (C) (C): a high-load range including full engine load

[0317] That is, on the operating range map 701, compared with the above-described operating range map 700, each of the low load range (A), the medium load range (B), and the high load range (C) is expanded in the engine rotation direction.

[0318] In the engine 1 applying the operating range map 701, particularly in order to expand the middle load range (B) in which the SPCCI combustion is performed, the swirl flow stronger than a given strength is generated in the combustion chamber 17.

[0319] For example, the ECU 10 adjusts the opening of the SCV 56 as needed. When the operating condition of the engine 1 is within the mid-load range (B), the SCV 56 is adjusted to be substantially fully closed or to have a given narrow opening. As a result, a swirl flow with the given strength is formed in the combustion chamber 17. At the time of reaching ignition from the intake stroke, the swirl ratio may be 4 or higher. Furthermore, the swirl ratio may be set below 4, for example, in a range from about 1.5 to about 3.

[0320] Also, in the engine 1 using the operating range map 701, the SI ratio is preferably adjusted similarly to the engine 1 using the operating range map 700. That is, as in Fig. As shown in Figure 11, within the second mid-load segment (B2), where boosting is not performed, the SI ratio increases linearly with increasing engine speed. As a result, SPCCI combustion is performed in a stable manner even within the high-speed range, while NVH, which can become a problem, is kept below the permissible level.

[0321] Further, in the engine 1 using the operating range map 701, the SI ratio is increased at a high engine speed by the first to fourth methods, similarly to the engine 1 using the operating range map 700. (Third operating range map)

[0322] Fig. 7C shows a third configuration example of the operating range map of engine 1 (third operating range map 702). The operating range map 702 of engine 1 is divided into the following four ranges. (1)-1: a low-load range that includes idle operation and extends into low and medium engine speed ranges (1)-2: a medium load range in which the engine load is higher than the low load range and which extends into the low and medium engine speed ranges (2): a high-load medium-speed range in which the engine load is higher than the medium-load range and which is located in the high-load range including or containing the full engine load (3): a high-load low-speed range located in the high-load range and in which the engine speed is lower than the high-load medium-speed range (4): a high-speed range in which the engine speed is higher than the low-load range, the medium-load range, the high-load medium-speed range and the high-load low-speed range

[0323] Here, the low-speed range, the medium-speed range, and the high-speed range can be defined by dividing the entire operating range of the engine 1 substantially equally in the engine speed direction into three ranges: the low-speed range, the medium-speed range, and the high-speed range. On the operating range map 702, the engine speed lower than the speed N1 is defined as low, the engine speed higher than a speed N2 is defined as high, and the engine speed between the engine speeds N1 and N2 is defined as medium.

[0324] For example, speed N1 may be approximately 1,200 rpm and speed N2 may be approximately 4,000 rpm. Note that speed N1 and speed N2 may be the same or different between operating range map 700 and operating range map 702.

[0325] On the operating range map 702, mainly to improve fuel efficiency and exhaust gas performance, the engine 1 performs SPCCI combustion within the low-load range (1)-1, the medium-load range (1)-2, and the high-load medium-speed range (2). The operating range map 702 differs from the operating range map 701 in that SPCCI combustion is also performed when the engine 1 operates at a low load and when the engine 1 operates at a high load. Furthermore, the engine 1 performs spark-ignition combustion within the other ranges, specifically, the high-load low-speed range (3) and the high-speed range (4). Hereinafter, the operation of the engine 1 within each range will be described in detail with reference to the fuel injection timing and the ignition timing shown in Fig. 18 are shown. (Low load range (1)-1)

[0326] When engine 1 operates within the low load range (1)-1, engine 1 performs SPCCI combustion.

[0327] The reference number 601 of Fig. 18 shows an example of the fuel injection timing (reference numerals 6011 and 6012) and the ignition timing (reference numeral 6013) and a combustion waveform (ie, a waveform indicating a change in heat generation rate with respect to the crank angle; reference numeral 6014) when the engine 1 operates in the operating state of reference numeral 601 within the low load range (1)-1 of the engine 1.

[0328] When the engine 1 operates within the low load range (1)-1, the swirl flow with the given strength is formed in the combustion chamber 17 (see Fig. 14B). The swirl ratio when the engine 1 operates within the low-load range (1)-1 is approximately 2 and is set within a range of less than 4, for example, within the range of 1.5 to 3. Note that a strong swirl flow can be formed by setting the swirl ratio to 4 or higher.

[0329] The swirl flow is strong in an outer peripheral portion of the combustion chamber 17 and weak in a central portion. The SCV 56 is substantially completely closed or has a given narrow opening. As described above, since the intake port 18 is a tumble port, an oblique swirl flow with a tumble component and a swirl component is formed in the combustion chamber 17.

[0330] When the engine 1 operates within the low-load range (1)-1, the air-fuel ratio (A / F) of the gas mixture is leaner than the theoretical air-fuel ratio in the entire combustion chamber 17. That is, the excess air ratio λ of the gas mixture exceeds 1 in the entire combustion chamber 17. In particular, the A / F of the gas mixture in the entire combustion chamber 17 is about 30:1 or higher. In this way, the formation of raw NO x reduced and exhaust performance is improved.

[0331] When the engine 1 operates within the low load range (1)-1, the EGR system 55 introduces the EGR gas into the combustion chamber 17 as needed. For example, the internal EGR gas is introduced according to the setting of the positive overlap period.

[0332] When the engine 1 operates within the low-load range (1)-1, the gas mixture is stratified between the central portion and the outer peripheral portion of the combustion chamber 17. The central portion of the combustion chamber 17 is the portion where the spark plug 25 is disposed, and the outer peripheral portion is the portion around the central portion and in contact with a liner of the cylinder 11. The central portion of the combustion chamber 17 may be defined as a portion where the swirl flow is weak, and the outer peripheral portion may be defined as a portion where the swirl flow is strong.

[0333] The fuel concentration of the gas mixture in the central portion is higher than that in the outer peripheral portion. For example, the A / F of the gas mixture in the central portion is between about 20:1 and about 30:1, and the A / F of the gas mixture in the outer peripheral portion is about 35:1 or higher.

[0334] When the engine 1 operates within the low-load range (1)-1, the injector 6 generally injects fuel into the combustion chamber 17 by dividing it into a plurality of injections during the compression stroke. The gas mixture is stratified in the central portion and the outer peripheral portion of the combustion chamber 17 by the divided fuel injections and the strong swirl flow in the combustion chamber 17.

[0335] After the fuel injection is completed, the spark plug 25 ignites the gas mixture in the central portion of the combustion chamber 17 at the given timing before the CTDC (see reference numeral 6013). Since the fuel concentration of the gas mixture in the central portion is relatively high, ignitability improves, and SI combustion by flame propagation stabilizes. By stabilizing SI combustion, CI combustion starts at the appropriate timing. The controllability of CI combustion improves in SPCCI combustion. When the engine 1 operates within the low-load range (1)-1, both the reduction in combustion noise generation and the improvement in fuel efficiency by shortening the combustion duration are achieved. (Medium load range (1)-2)

[0336] Even when engine 1 operates within the medium-load range (1)-2, engine 1 performs SPCCI combustion similarly to the low-load range (1)-1. The medium-load range (1)-2 corresponds to the medium-load range (B) on the operating range map 701.

[0337] The reference number 602 of Fig. 18 shows an example of the fuel injection timing (reference numerals 6021 and 6022) and the ignition timing (reference numeral 6023) and a combustion waveform (reference numeral 6024) when the engine 1 operates in the operating state of reference numeral 602 within the middle load range (1)-2 of the engine 1.

[0338] The EGR system 55 introduces the EGR gas into the combustion chamber 17 when the operating state of the engine 1 is within the medium load range (1)-2. For example, the internal EGR gas is introduced according to the setting of the positive overlap period.

[0339] Even when the engine 1 operates within the medium load range (1)-2, similar to the low load range (1)-1, a swirl flow with a given strength, where the swirl ratio is approximately 2, is formed in the combustion chamber 17. The SCV 56 is not completely closed but has a given narrow opening. By forming the swirl flow, the residual gas accumulated in the cavity 31 is discharged from the cavity 31. As a result, the G / F of the gas mixture in the SI section near the spark plug 25 can be varied from the G / F of the gas mixture in the CI section around the SI section. Therefore, the SPCCI combustion is stabilized by setting the total G / F in the entire combustion chamber 17 to between 18 and 50, as described above.

[0340] Since the kinetic energy of turbulence in the combustion chamber 17 increases due to the formation of swirl flow when the engine 1 operates within the medium load range (1)-2, the SI combustion flame propagates immediately, and the SI combustion is stabilized. The controllability of the CI combustion improves by stabilizing the SI combustion. By appropriately adjusting the timing of the CI combustion in the SPCCI combustion, the generation of combustion noise is reduced and fuel efficiency is improved. Furthermore, the torque variation between cycles is reduced.

[0341] When the engine 1 operates within the medium load range (1)-2, the air-fuel ratio (A / F) of the gas mixture is substantially at the theoretical air-fuel ratio (A / F = about 14.7:1) throughout the combustion chamber 17. The three-way catalyst purifies the exhaust gas discharged from the combustion chamber 17. Thus, the exhaust performance of the engine 1 becomes good. The A / F of the gas mixture can be adjusted to remain within the purification window of the three-way catalyst. Therefore, the excess air ratio λ of the gas mixture can be set to about 1.0 ± about 0.2.

[0342] When the engine 1 operates within the medium load range (1)-2, the injector 6 performs fuel injection on the intake stroke (reference numeral 6021) and fuel injection on the compression stroke (reference numeral 6022). By performing the first injection 6021 on the intake stroke, the fuel is distributed substantially evenly in the combustion chamber 17. By performing the second injection 6022 on the compression stroke, the temperature in the combustion chamber 17 is reduced by the latent heat of vaporization of the fuel. The gas mixture containing the fuel injected in the first injection 6021 is prevented from causing preignition.

[0343] When the injector 6 performs the first injection 6021 during the intake stroke and the second injection 6022 during the compression stroke, the gas mixture with the excess air ratio λ of 1.0 ± 0.2 is formed in the combustion chamber 17 as a whole. Since the fuel concentration of the gas mixture is substantially homogeneous, the improvement in fuel efficiency is achieved by reducing the loss of unburned fuel and the improvement in exhaust performance by preventing smoke generation. The excess air ratio λ is preferably 1.0 to 1.2. Furthermore, the total G / F in the entire combustion chamber 17 is about 18:1 to about 50:1, and the G / F of the SI portion near the spark plug 25 is about 14:1 to about 22:1.

[0344] By igniting the gas mixture at the specified timing before the CTDC (reference numeral 6023), the gas mixture burns by flame propagation. After combustion by flame propagation begins, the unburned gas mixture self-ignites at the desired timing, causing CI combustion. The fuel injected during the second-stage injection primarily causes SI combustion. The fuel injected during the first-stage injection primarily causes CI combustion. By maintaining the total G / F in the entire combustion chamber 17 between 18:1 and 50:1 and the G / F of the SI portion near the spark plug 25 between 14:1 and 22:1, SPCCI combustion is stabilized.

[0345] According to the illustration in Fig. 7C is a range in which the booster 44 is turned off (see S / C OFF), a part of the low-load range (1)-1 and a part of the medium-load range (1)-2. In detail, the booster 44 is turned off within the low-speed segment of the low-load range (1)-1. Within a high-speed segment of the low-load range (1)-1, the booster 44 is turned on to increase the boosting pressure to ensure a required intake charge amount corresponding to the engine speed, which is high. Furthermore, the booster 44 is turned off within a low-load low-speed segment of the medium-load range (1)-2. Within a high-load segment of the medium-load range (1)-2, the booster 44 is turned on to ensure the required intake charge amount corresponding to the engine speed, which is high.Within a high-speed segment of the medium-load range (1)-2, the booster 44 is switched on to ensure the required intake charge amount according to the engine speed, which is high.

[0346] It should be noted that within the high-load medium-speed range (2), the high-load low-speed range (3) and the high-speed range (4), the booster 44 is switched on across the ranges. (High-load medium speed range (2))

[0347] Even when the engine 1 operates within the high-load medium-speed range (2), the engine 1 performs the SPCCI combustion similar to the low-load range (1)-1 and the medium-load range (1)-2.

[0348] The reference number 603 of Fig. Fig. 18 shows an example of the fuel injection timing (reference numerals 6031 and 6032) and the ignition timing (reference numeral 6033) and a combustion waveform (reference numeral 6034) when the engine 1 operates in the operating state of reference numeral 603 within the high-load medium speed range (2) of the engine 1. Furthermore, reference numeral 604 of Fig. 18 shows an example of the fuel injection timing (reference numeral 6041) and the ignition timing (reference numeral 6042) and a combustion waveform (reference numeral 6043) when the engine speed is higher than in the operating state of reference numeral 603.

[0349] The EGR system 55 introduces EGR gas (e.g., external EGR gas) into the combustion chamber 17 when the operating condition of the engine 1 is within the high-load / mid-speed range (2). The engine 1 reduces the amount of EGR gas as the engine load increases. At full engine load, the EGR gas can be set to approximately zero.

[0350] Even when the engine 1 operates within the high-load medium-speed range (2) similarly to the low-load range (1)-1, a swirl flow with the given strength is formed in the combustion chamber 17, with the swirl ratio being below 4. A strong swirl flow can be formed by setting the swirl ratio to 4 or higher. The SCV 56 is controlled to have a given narrow opening without being substantially completely closed.

[0351] When the engine 1 operates within the high-load medium-speed range (2), the air-fuel ratio (A / F) of the gas mixture is at or richer than the theoretical air-fuel ratio in the entire combustion chamber 17 (ie, the excess air ratio λ of the gas mixture is less than or equal to 1 (λ<=1)).

[0352] When engine 1 operates on the low-speed side of the high-load medium-speed range (2), injector 6 injects fuel during the intake stroke (reference numeral 6031) and injects fuel during a final stage of the compression stroke (reference numeral 6032). The final stage of the compression stroke can be defined by equally dividing the compression stroke into three stages: an initial stage, an intermediate stage, and the final stage.

[0353] First-stage injection 6031, which begins on the intake stroke, can start fuel injection in the early half of the intake stroke. The early half of the intake stroke can be defined by equally dividing the intake stroke into two parts, the first half and the second half. For example, first-stage injection can start fuel injection at approximately 280°CA before TDC.

[0354] When the injection of the first-stage injection 6031 is started in the early half of the intake stroke, the fuel spray, although not shown, hits an opening edge of the cavity 31, so that part of the fuel enters the squish area 171 of the combustion chamber 17, and the rest of the fuel enters the portion inside the cavity 31. The swirl flow is strong in the outer peripheral portion of the combustion chamber 17 and weak in the central portion. Therefore, the part of the fuel that has entered the squish area 171 joins the swirl flow, and the rest of the fuel that has entered the portion inside the cavity 31 joins the inside of the swirl flow. The fuel joins the swirl flow, remains in the swirl flow during the intake stroke and the compression stroke, and forms the gas mixture for CI combustion in the outer peripheral portion of the combustion chamber 17.The fuel entering the inside of the swirl flow also remains on the inside of the swirl flow during the intake stroke and the compression stroke and forms the gas mixture for the SI combustion in the middle section of the combustion chamber 17.

[0355] When the engine 1 operates within the high-load medium-speed range (2), the excess air ratio λ of the gas mixture in the central portion where the spark plug 25 is arranged is preferably about 1 or less, and the excess air ratio λ of the gas mixture in the outer peripheral portion is about 1 or less, preferably below 1. The air-fuel ratio (A / F) of the gas mixture in the central portion may be, for example, between about 13 and about the theoretical air-fuel ratio (14.7:1). The air-fuel ratio of the gas mixture in the central portion may be leaner than the theoretical air-fuel ratio. Further, the air-fuel ratio of the gas mixture in the outer peripheral portion may be, for example, between about 11:1 and about the theoretical air-fuel ratio, preferably between about 11:1 and about 12:1.Since the amount of fuel in the gas mixture in the outer peripheral portion increases when the excess air ratio λ of the outer peripheral portion of the combustion chamber 17 is set below 1, the temperature is reduced by the latent heat of vaporization of the fuel. The air-fuel ratio of the gas mixture in the entire combustion chamber 17 may be between about 12.5:1 and about the theoretical air-fuel ratio, preferably between about 12.5:1 and about 13:1.

[0356] The second-stage injection 6032, which is performed in the final stage of the compression stroke, can start fuel injection at 10°CA before TDC. By performing the second-stage injection immediately before TDC, the temperature in the combustion chamber 17 is lowered by the latent heat of vaporization of the fuel. Although a low-temperature oxidation reaction of the fuel injected in the first-stage injection 6031 progresses during the compression stroke and transitions to a high-temperature oxidation reaction before TDC, by performing the second-stage injection 6032 immediately before TDC to thus lower the temperature inside the combustion chamber, the transition from the low-temperature oxidation reaction to the high-temperature oxidation reaction is avoided, and preignition is prevented.It should be noted that the ratio between the injection quantity of the first stage injection and the injection quantity of the second stage injection may be, for example, approximately 95:5.

[0357] The spark plug 25 ignites the gas mixture in the central portion of the combustion chamber 17 near CTDC (reference numeral 6033). The spark plug 25 ignites, for example, after CTDC. Since the spark plug 25 is located in the central portion of the combustion chamber 17, the ignition of the spark plug 25 causes the gas mixture in the central portion to begin SI combustion through flame propagation.

[0358] Within the high-load range, the fuel injection amount increases, as does the temperature of the combustion chamber 17, so CI combustion is likely to start early. In other words, pre-ignition of the gas mixture is likely to occur within the high-load range. However, since the temperature of the outer peripheral portion of the combustion chamber 17 is lowered by the latent heat of vaporization of the fuel as described above, CI combustion is prevented from starting immediately after the spark ignition of the gas mixture.

[0359] As described above, when the spark plug 25 ignites the gas mixture in the central portion, the combustion speed increases, and the SI combustion is stabilized by high turbulence kinetic energy, and the SI combustion flame propagates in the circumferential direction along the strong swirl flow within the combustion chamber 17. Then, at a given position of the outer peripheral portion of the combustion chamber 17 in the circumferential direction, the unburned gas mixture ignites itself, and the CI combustion begins.

[0360] In the concept of this SPCCI combustion, by combining the stratification of the gas mixture in the combustion chamber 17 and the swirl flow in the combustion chamber 17, the SI combustion is sufficiently carried out until the CI combustion begins. As a result, the generation of combustion noise is reduced, and since the combustion temperature does not become excessively high, the generation of NO is also reduced. x reduced. Furthermore, the torque change between cycles is reduced.

[0361] Furthermore, since the temperature in the outer peripheral portion is low, CI combustion becomes slower and the generation of combustion noise is reduced. Furthermore, since the combustion duration is shortened by CI combustion, torque and thermal efficiency are improved within the high-load range. Thus, by performing SPCCI combustion within the high engine load range, the fuel efficiency of the engine 1 is improved while combustion noise is avoided.

[0362] When the engine 1 operates on the high-speed side of the high-load medium-speed range (2), the injector 6 starts the fuel injection on the intake stroke (reference numeral 6041).

[0363] The first-stage injection 6041, which starts on the intake stroke, may start fuel injection in the early half of the intake stroke, similar to the description above. For example, the first-stage injection 6041 may start fuel injection at approximately 280°CA before TDC. The first-stage injection may continue throughout the intake stroke and end on the compression stroke. By setting the injection start time of the first-stage injection 6041 in the early half of the intake stroke, the gas mixture for CI combustion is formed in the outer peripheral portion of the combustion chamber 17, and the gas mixture for SI combustion is formed in the central portion of the combustion chamber 17.Similar to the above description, the excess air ratio λ of the gas mixture in the central portion where the spark plug 25 is arranged is preferably about 1 or less, and the excess air ratio λ of the gas mixture in the outer peripheral portion is about 1 or less, preferably below 1. The air-fuel ratio (A / F) of the gas mixture in the central portion may be, for example, between about 13:1 and about the theoretical air-fuel ratio (14.7:1). The air-fuel ratio of the gas mixture in the central portion may be leaner than the theoretical air-fuel ratio. Further, the air-fuel ratio of the gas mixture in the outer peripheral portion may be, for example, between about 11:1 and about the theoretical air-fuel ratio, preferably between about 11:1 and about 12:1.The air-fuel ratio of the gas mixture in the entire combustion chamber 17 may be between about 12.5:1 and about the theoretical air-fuel ratio, preferably between about 12.5:1 and about 13:1.

[0364] As the engine speed increases, the time for the fuel injected during the first-stage injection 6041 to react becomes shorter. Therefore, the second-stage injection can be omitted to suppress the oxidation reaction of the gas mixture.

[0365] Spark plug 25 ignites the gas mixture in the central portion of combustion chamber 17 near CTDC (reference numeral 6042). Spark plug 25 ignites, for example, after CTDC.

[0366] As described above, by layering or stratifying the gas mixture within the high-load medium-speed range (2), the combustion noise is reduced and the SPCCI combustion is stabilized. (High load low speed range (3))

[0367] When engine 1 operates within the high-load low-speed range (3), engine 1 performs SI combustion instead of SPCCI combustion. The high-load low-speed range (3) corresponds to the first high-load segment (C1) on the operating range map 701.

[0368] Reference numeral 605 of Fig. 18 shows an example of the fuel injection timings (reference numerals 6051 and 6052) and the ignition timing (reference numeral 6053) and a combustion waveform (reference numeral 6054) when the engine 1 operates in the operating state of reference numeral 605 within the high-load low-speed range (3) of the engine 1.

[0369] The EGR system 55 introduces EGR gas into the combustion chamber 17 when the operating condition of the engine 1 is within the high-load / low-speed range (3). The engine 1 reduces the amount of EGR gas as the engine load increases. At full load, the EGR gas can be set to approximately zero.

[0370] When the engine 1 operates within the high-load low-speed range (3), the air-fuel ratio (A / F) of the gas mixture is substantially at the theoretical air-fuel ratio (A / F = about 14.7:1) throughout the combustion chamber 17. The A / F of the gas mixture can be adjusted to remain within the purification window of the three-way catalyst. Therefore, the excess air ratio λ of the gas mixture can be adjusted to about 1.0 ± about 0.2. By adjusting the air-fuel ratio of the gas mixture substantially at the theoretical air-fuel ratio, fuel efficiency improves within the high-load low-speed range (3).When the engine 1 operates within the high-load low-speed range (3), the fuel concentration of the gas mixture in the entire combustion chamber 17 can be adjusted so that the excess air ratio λ is about 1 or less and equal to or higher than the excess air ratio λ within the high-load medium-speed range (2), preferably higher than the excess air ratio λ within the high-load medium-speed range (2).

[0371] On the operating range map 702, when the engine 1 operates within the high-load low-speed range (3), the injector 6 injects fuel into the combustion chamber 17 at the timing of the intake stroke and in the lag period from the final stage of the compression stroke to the early stage of the expansion stroke (reference numerals 6051 and 6052). By injecting the fuel in two injections, the amount of fuel injected in the lag period is reduced. By injecting the fuel during the intake stroke (reference numeral 6051), the formation period of the gas mixture is sufficiently ensured. In addition, by injecting the fuel in the lag period (reference numeral 6052), the flow in the combustion chamber 17 immediately before ignition is increased, which is beneficial for stabilizing SI combustion.

[0372] After the fuel is injected, spark plug 25 ignites the gas mixture at a time near CTDC (reference numeral 6053). For example, spark plug 25 ignites after CTDC. The gas mixture causes SI combustion on the expansion stroke. Since SI combustion begins on the expansion stroke, CI combustion does not begin.

[0373] When engine 1 operates within the high-load low-speed range (3), the swirl flow is made weaker than when it operates within the high-load medium-speed range (2). When engine 1 operates within the high-load low-speed range (3), the opening of the SCV 56 is larger than when it operates within the high-load medium-speed range (2). For example, the opening of the SCV 56 may be approximately 50% (i.e., half-open).

[0374] The axes of the nozzle openings of the injector 6 do not overlap circumferentially with the spark plug 25. The fuel injected from the nozzle openings flows circumferentially due to the swirl flow in the combustion chamber 17. This swirl flow immediately transports the fuel to the vicinity of the spark plug 25. The fuel is vaporized while being transported to the vicinity of the spark plug 25.

[0375] On the other hand, if the swirl flow is excessively strong, the fuel flows in the circumferential direction and moves away from the spark plug 25, and the fuel cannot be promptly delivered to the vicinity of the spark plug 25. For this reason, when the engine 1 operates within the high-speed / low-speed range (3), the swirl flow is made weaker than when operating within the high-load / medium-speed range (2). As a result, the fuel is promptly delivered to the vicinity of the spark plug 25, thereby improving the ignitability of the gas mixture and stabilizing the SI combustion. (High speed range (4))

[0376] When the engine speed is high, the time required for the crank angle to change becomes shorter by 1°. Therefore, for example, on the high-speed side of the high-load range, it is difficult to stratify the gas mixture in the combustion chamber 17 as described above. As the engine speed increases, it becomes difficult to perform the SPCCI combustion described above.

[0377] When engine 1 operates within the high-speed range (4), engine 1 performs SI combustion instead of SPCCI combustion. Note that the high-speed range (4) extends across the entire load direction from low load to high load.

[0378] The reference number 606 of Fig. 18 shows an example of the fuel injection timing (reference numeral 6061) and the ignition timing (reference numeral 6062) and a combustion waveform (reference numeral 6063) when the engine 1 operates in the operating state of reference numeral 606 within the high speed range (4).

[0379] The EGR system 55 introduces EGR gas into the combustion chamber 17 when the operating state of the engine 1 is within the high-speed range (4). The engine 1 reduces the amount of EGR gas as the engine load increases. At full load, the EGR gas can be set to approximately zero.

[0380] When engine 1 operates within the high-speed range (4), it opens the SCV 56 substantially completely. No swirl flow is generated in the combustion chamber 17, and only tumble flow is generated. By substantially completely opening the SCV 56, charging efficiency in the high-speed range (4) is improved and pumping loss is reduced.

[0381] When the engine 1 operates within the high-speed range (4), the air-fuel ratio (A / F) of the gas mixture is basically substantially at the theoretical air-fuel ratio (A / F = about 14.7:1) throughout the combustion chamber 17. The excess air ratio λ of the gas mixture can be set to about 1.0 ± about 0.2. Note that within the high-load side of the high-speed range (4), including or including the full load, the excess air ratio λ of the gas mixture can be less than 1.

[0382] When the engine 1 is operating within the high-speed range (4), the injector 6 starts fuel injection on the intake stroke (reference numeral 6061). The injector 6 injects all the fuel for one combustion cycle in one go. By starting fuel injection on the intake stroke, a homogeneous or substantially homogeneous gas mixture is formed in the combustion chamber 17. Furthermore, when the engine speed is high, since the evaporation period of the fuel is ensured as long as possible, the loss of unburned fuel and the generation of soot are reduced.

[0383] After the fuel injection is completed, the spark plug 25 ignites the gas mixture at an appropriate time before CTDC (reference numeral 6062).

[0384] Also, in the engine 1 using the operating range map 702, the SI ratio is preferably set similarly to the engine 1 using the operating range map 700. That is, as in Fig. As shown in Figure 11, within the low-load range (1)-1, the medium-load range (1)-2, and the high-load medium-speed range (2) where SPCCI combustion is performed, the SI ratio increases linearly with increasing engine speed. As a result, SPCCI combustion is performed stably, even within the high-speed range, while NVH, which could become a problem, is kept below the allowable value.

[0385] Further, in the engine 1 using the operating range map 702, the SI ratio is increased at a high engine speed by the first to fourth methods, similarly to the engine 1 using the operating range map 700. (combustion waveform)

[0386] Fig. 19 and Fig. 20 show combustion waveforms in operating states of the engine 1 using the operating range map 702. Fig. 19 is a short graph corresponding to the third operating range map 702 shown in Fig. 7C is shown.

[0387] The combustion waveform indicates the change in heat generation rate with respect to the change in crank angle. W1 to W12 correspond to the Fig. 19 shown respective operating states.

[0388] As described above, within the range where SPCCI combustion is performed, the SI ratio increases as the engine speed increases. When comparing the waveforms of W2, W7, and W10, the peak of SI combustion gradually increases in the order of W2, W7, and W10. As a result, the peak of CI combustion gradually decreases in the order of W2, W7, and W10. Thus, the generation of combustion noise is reduced as the engine speed increases. Note that within the SPCCI range, the similar trend is shown in waveforms W3, W8, and W11 and waveforms W4, W9, and W12, which are for lower engine loads than W2, W7, and W10. (Other embodiments)

[0389] Note that the control of engine 1 performed by ECU 10 is not limited to being based on the combustion model described above. For example, although the internal EGR system of engine 1 described above changes the internal EGR ratio by changing the negative overlap period, it can also be changed by opening the exhaust valve during the intake stroke.

[0390] The third and fourth methods are not limited to supplementing the first or second methods. They can be independent of the first and second methods depending on the specifications. This means that the combination of the first to fourth methods is freely selectable.

[0391] Furthermore, instead of performing SI combustion within the high-speed range, engine 1 can perform SPCCI combustion across the entire engine speed range. Furthermore, engine 1 can perform SPCCI combustion across the entire operating range.

[0392] Furthermore, the prior art disclosed herein is not limited to being applied to the engine 1 having the above configuration. The configuration of the engine 1 may take various configurations. List of reference symbols 1 engine 10 ECUs (controllers) 17 Combustion chamber 23 electrically operated intake S-VT (internal EGR system) 24 electrically operated exhaust S-VT (internal EGR system) 25 Spark plug 3 pistons 31 cavity 54 EGR valve (external EGR system) 55 EGR (Exhaust Gas Recirculation) System 56 Swirl control or regulating valve (inlet flow control or regulating device) 6 Injector 102a SI ratio change modulus (heat ratio change modulus) SW6 pressure sensor

Claims

[1] Internal combustion engine, comprising: a combustion chamber (17) configured to cause combustion of a gas mixture within the combustion chamber (17); a spark plug (25) arranged to be substantially aligned into the combustion chamber (17) and configured to ignite the gas mixture within the combustion chamber (17); and a controller (10) operatively connected to the spark plug (25), wherein the combustion is carried out in at least one given mode (SPCCI (SPark Controlled Compression Ignition) combustion) in which spark ignition combustion (SI combustion) and compression ignition combustion (CI combustion) are combined, in which, after the spark plug (25) has ignited the gas mixture to start combustion, unburned gas mixture burns by self-ignition, wherein the controller (10) comprises a heat quantity ratio changing module (102a) configured to change, according to an operating state of the engine, a heat quantity ratio as an index concerning a ratio of the heat quantity generated by the SI combustion with respect to a total heat quantity generated by the SPCCI combustion, and wherein the controller (10) causes the heat quantity ratio changing module (102a) to increase the heat quantity ratio at a high engine speed compared to a low engine speed. [2] The internal combustion engine according to claim 1, further comprising a pressure sensor (SW6) configured to detect a pressure in the combustion chamber (17), wherein the controller (10) is configured to set a target heat quantity ratio, receive a detection signal from the pressure sensor (SW6), calculate the heat quantity ratio based on a pressure waveform caused by the combustion of the gas mixture, and, when the calculated heat quantity ratio is different from the target heat quantity ratio, change the heat quantity ratio to approach the target heat quantity ratio. [3] The internal combustion engine according to claim 1 or 2, further comprising an internal exhaust gas recirculation (EGR) system (55) configured to change an internal EGR ratio, which is a ratio of an amount of internal EGR gas contained in the gas mixture within the combustion chamber (17), wherein the controller (10) outputs a control signal to the internal EGR system (55) to increase the internal EGR ratio to increase the heat amount ratio. [4] An internal combustion engine according to claim 3, wherein the internal EGR gas is introduced into the combustion chamber (17) by overlapping an open period of an intake valve (21) with an open period of an exhaust valve (22). [5] An internal combustion engine according to any one of the preceding claims, further comprising an external EGR system (54) configured to change an external EGR ratio, which is a ratio of an amount of external EGR gas contained in the gas mixture within the combustion chamber (17), wherein the controller (10) outputs a control signal to the external EGR system (54) to decrease the external EGR ratio so as to increase the heat amount ratio. [6] An internal combustion engine according to any one of the preceding claims, wherein a swirl flow is formed in the combustion mode when the combustion is carried out in the given mode. [7] An internal combustion engine according to claim 6, wherein a swirl ratio of the swirl flow is less than 4. [8] An internal combustion engine according to claim 6 or 7, further comprising a swirl control valve (56) configured to change a strength of the swirl flow by controlling an opening, wherein the strength of the swirl flow increases when the opening of the swirl control or regulating valve is reduced, and wherein the controller (10) controls the swirl control valve (56) to have a given narrow opening without completely closing the swirl control valve (56). [9] An internal combustion engine according to any one of the preceding claims, wherein the controller (10) outputs the control signal to the spark plug (25) to advance an ignition timing to increase the heat quantity ratio. [10] Internal combustion engine, comprising: a combustion chamber (17) configured to cause combustion of gas mixture within the combustion chamber (17); a spark plug (25) arranged to be substantially aligned into the combustion chamber (17) and configured to ignite the gas mixture within the combustion chamber (17); an EGR system (55) configured to change an EGR ratio, which is a ratio of an amount of EGR gas contained in the gas mixture within the combustion chamber (17); and a controller (10) comprising a processor operatively connected to the spark plug (25) and the EGR system (55) and configured to operate the engine (1) by outputting a control signal to the spark plug (25) and the EGR system (55), respectively, wherein the controller (10) outputs the control signal to the spark plug (25) at a given ignition timing to perform combustion in at least one given mode in which, after the spark plug (25) ignites the gas mixture to start combustion, unburned gas mixture in the combustion chamber burns by auto-ignition, and outputs the control signal to the EGR system (55) to make the EGR ratio higher at a high engine speed than at a low engine speed. [11] An internal combustion engine according to claim 10, wherein when the engine speed exceeds a given limit start speed, the controller (10) outputs the control signal to the EGR system (55) to limit the increase in the EGR ratio. [12] An internal combustion engine according to claim 11, wherein the limit start speed is set to be lower as the engine speed increases. [13] An internal combustion engine according to any one of the preceding claims 10 to 12, further comprising an intake flow control device (56) configured to change a flow of intake air introduced into the combustion chamber (17), wherein the controller (10) outputs a control signal to the intake flow control device (56) to increase the flow of intake air within an engine speed range exceeding the limit start speed. [14] An internal combustion engine according to any one of the preceding claims 10 to 13, wherein the controller (10) outputs a control signal to the spark plug (25) so that an ignition timing advances within an engine speed range exceeding the limit start speed. [15] An internal combustion engine according to any one of the preceding claims 10 to 14, wherein the EGR gas is an internal EGR gas introduced into the combustion chamber (17) by providing a negative overlap period. [16] Internal combustion engine according to claim 15, further comprising a boosting system (44) configured to boost gas to be introduced into the combustion chamber (17), wherein the controller (10) outputs a control signal to the boosting system (44) to perform boosting within a first range in which an engine load is higher than a given load, and not to perform boosting within a second range in which the engine load is below the given load, and wherein within the second range the controller (10) outputs the control signal to the EGR system (55) to increase the EGR ratio higher at a high engine speed than at a low engine speed. [17] An internal combustion engine according to any one of the preceding claims 10 to 16, wherein the EGR gas is an internal EGR gas introduced into the combustion chamber (17) by overlapping an open period of an intake valve (21) with an open period of an exhaust valve (22). [18] An internal combustion engine according to any one of the preceding claims 10 to 17, wherein the EGR ratio continuously increases without limitation until the engine speed becomes close to the highest speed within an operating range of the engine in which combustion of the given mode is performed. [19] Internal combustion engine according to one of the preceding claims 10 to 18, wherein a swirl flow is formed in the combustion chamber (17) when the combustion is carried out in the given mode. [20] Control or regulating device for an internal combustion engine which can be operated in at least one given mode (SPCCI (SPark Controlled Compression Ignition) combustion) in which spark ignition combustion (SI combustion) and compression ignition combustion (CI combustion) are combined, in which, after a spark plug (25) ignites the gas mixture to start the combustion, unburned gas mixture burns by self-ignition, wherein the control or regulating deviceA control device comprising a controller (10) having a heat quantity ratio changing module (102a) configured to change, according to an operating state of the engine, a heat quantity ratio as an index relating to a ratio of the heat quantity generated by the SI combustion with respect to a total heat quantity generated by the SPCCI combustion, and wherein the controller (10) causes the heat quantity ratio changing module (102a) to increase the heat quantity ratio at a high engine speed compared to a low engine speed. [21] The control device according to claim 20, further comprising a pressure sensor (SW6) configured to detect a pressure in the combustion chamber (17), wherein the controller (10) is configured to set a target heat quantity ratio, receive a detection signal from the pressure sensor (SW6), calculate the heat quantity ratio based on a pressure waveform caused by the combustion of the gas mixture, and, when the calculated heat quantity ratio is different from the target heat quantity ratio, change the heat quantity ratio to approach the target heat quantity ratio. [22] A control device for an internal combustion engine that can be operated in at least one given mode in which, after a spark plug (25) ignites the gas mixture to start combustion, unburned gas mixture burns by self-ignition, and that comprises an EGR system (55) configured to change an EGR ratio, which is a ratio of an amount of EGR gas contained within the gas mixture within a combustion chamber (17), wherein the control device further comprises a controller (10) operatively connected to the spark plug (25) and the EGR system (55) and configured to operate the engine by outputting a control signal to the spark plug (25) and the EGR system (55), respectively, wherein the controller (10) outputs the control signal to the spark plug (25) at a given ignition timing to perform combustion in the given mode, and the control orOutputs a control signal to the EGR system (55) to increase the EGR ratio higher at a high engine speed than at a low engine speed. [23] A control device according to claim 22, wherein, when the engine speed exceeds a given limit start speed, the controller (10) outputs the control signal to the EGR system (55) to limit the increase in the EGR ratio. [24] Method for controlling or regulating an internal combustion engine, comprising the steps: Operating the engine in at least one given mode (SPCCI (SPark Controlled Compression Ignition) combustion) in which spark ignition combustion (SI combustion) and compression ignition combustion (CI combustion) are combined, in which, after a spark plug (25) ignites the gas mixture to start combustion, unburned gas mixture burns by self-ignition, Changing, according to an operating condition of the engine, a heat quantity ratio as an index concerning a ratio of the heat quantity generated by the SI combustion with respect to a total heat quantity generated by the SPCCI combustion, and Increasing the heat quantity ratio at a high engine speed compared to a low engine speed. [25] Method for controlling or regulating an internal combustion engine, comprising the steps: Operating the engine in at least one given mode in which, after a spark plug (25) ignites the gas mixture to start combustion, unburned gas mixture burns by self-ignition, Changing an EGR ratio, which is a ratio of an amount of EGR gas contained in the gas mixture within a combustion chamber (17), Actuating the spark plug (25) at a given ignition point to perform combustion in the given mode, and Increasing the EGR ratio higher at a high engine speed than at a low engine speed. [26] A computer program product comprising computer-readable instructions which, when loaded and executed on a suitable system, can perform the steps of a method according to claim 24 or 25.

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