Combustion control system for diesel engine

A combustion control system for a diesel engine provided with a fuel injection system and a sensor for detecting an ignition timing. The combustion control system comprises a control unit which is configured to control the ignition timing to be restored to a target value when the ignition timing shifts from the target value owing to circumferential factors such as cetane number of fuel. Preferably, the ignition timing is controlled to generally coincide with or be retarded relative to a fuel injection termination timing. This regulates an ignition delay period to be longer than a fuel injection period in a predetermined engine operating region, thereby effectively accomplishing so-called low temperature premix combustion.

The contents of Japanese Patent Application No. 9-266358, with a filing
 date of Sep. 30, 1997 in Japan, are hereby incorporated by reference.
 BACKGROUND OF THE INVENTION
 1. Field of the Invention
 This invention relates to improvements in a combustion control system for a
 diesel engine, and more particularly to the combustion control system for
 effecting a low temperature premix combustion in a combustion chamber of
 the engine so as to improve exhaust emission performance and combustion
 noise performance of the engine.
 2. Description of the Prior Art
 Hitherto a variety of techniques for improving exhaust gas performance and
 combustion noise performance of a diesel engine has been proposed and put
 into practical use. One of them is disclosed in Japanese Patent
 Provisional Publication No. 8-86251 and configured as follows: A fuel
 injection timing is retarded to a timing after top dead center on
 compression stroke thereby prolonging a so-called ignition delay period of
 between a fuel injection initiation timing and an ignition timing of
 injected fuel. During this prolonged ignition delay period, a pre-mixture
 of air and fuel is formed in the combustion chamber by lowering an oxygen
 concentration owing to exhaust gas recirculation and by controlling gas
 flow in the combustion chamber, thus accomplishing a so-called low
 temperature premix combustion.
 SUMMARY OF THE INVENTION
 Now, the low temperature premix combustion aims at the fact that fuel is
 dispersed around oxygen as much as possible before initiation of ignition.
 In this regard, the latest experiments have revealed that the most
 important condition for effectively realizing the low temperature premix
 combustion is to terminate fuel injection during the ignition delay period
 (i.e., the ignition delay period is longer than a fuel injection period
 for which fuel is being injected), in addition to promoted dispersion of
 fuel under a gas flow control. Accordingly, if the ignition delay period
 is changed owing to difference in circumferential factors such as
 difference in cetane number of fuel so as to become shorter than the fuel
 injection period, the low temperature premix combustion cannot be
 accomplished and therefore exhaust gas purification performance inherent
 in the low temperature premix combustion cannot be obtained. For example,
 in case that the cetane number of fuel is high so that the ignition delay
 period is shorter than the fuel injection period, the low temperature
 premix combustion cannot be realized and therefore diffusive combustion is
 mainly made in the combustion stroke thereby increasing smoke in exhaust
 gas.
 In case that the cetane number of fuel is low, the ignition delay period is
 prolonged and therefore such a low cetane number is preferable for the low
 temperature premix combustion. However, if the ignition timing of fuel is
 further retarded owing to use of low cetane number fuel upon having been
 originally retarded in the low temperature premix combustion, fuel economy
 is degraded owing to lowering in combustion pressure in a high EGR rate
 operating range while causing misfire thereby increasing unburned fuel
 emission. Additionally, this increases combustion noise in a low EGR rate
 operating range of the diesel engine.
 An object of the present invention is to provide an improved combustion
 control system for a diesel engine, which can overcome drawbacks
 encountered in conventional combustion control systems for diesel engines.
 Another object of the present invention is to provide an improved
 combustion control system for a diesel engine, which can effectively
 prevent performances of smoke, fuel economy and combustion noise from
 being degraded under the influence of circumferential factors such as
 cetane number of fuel.
 A further object of the present invention is to provide an improved
 combustion control system for a diesel engine, by which an ignition delay
 period is controlled to be longer than a fuel injection period thereby
 effectively accomplishing low temperature premix combustion.
 A still further object of the present invention is to provide an improved
 combustion control system for a diesel engine, which can accomplish such a
 control as to cause the relationship between an ignition delay period and
 a fuel injection period to fall into a predetermined state when an
 ignition timing shifts from a target value.
 A first aspect of the present invention resides in a combustion control
 system for a diesel engine, comprising a section for changing an ignition
 timing. A section is provided for changing a fuel injection period. A
 section is provided for detecting the ignition timing. Additionally, a
 section is provided for controlling a relationship between an ignition
 delay period and the fuel injection period to fall into a predetermined
 state when the detected ignition timing shifts from a target value.
 Accordingly, even if the ignition timing is changed owing to
 circumferential factors such as cetane number of fuel, smoke in exhaust
 gas, fuel consumption and combustion noise can be effectively prevented
 from being degraded.
 A second aspect of the present invention resides, as shown in FIG. 14, in a
 combustion control system for a diesel engine, comprising a section 81 for
 detecting an ignition timing. Additionally, a section 82 is provided for
 controlling the detected ignition timing to generally coincide with or be
 retarded relative to a fuel injection termination timing in a
 predetermined engine operating region. Accordingly, the ignition timing is
 necessarily controlled to generally coincide with or retarded relative to
 the fuel injection termination timing. As result, the ignition delay
 period is controlled to be longer than the fuel injection period, thereby
 effectively accomplishing the low temperature premix combustion. This can
 effectively prevent smoke in exhaust gas and combustion noise from
 increasing even if circumferential factors such as cetane number of fuel
 are changed.
 A third aspect of the present invention resides, as shown in FIG. 15, in a
 combustion control system for a diesel engine, comprising a section 81 for
 detecting an ignition timing. A section 91 is provided for judging that
 the detected ignition timing is retarded relative to a target value. A
 section 92 is provided for controlling the detected ignition timing so as
 to be restored to the target value when the detected ignition timing is
 retarded relative to the target value. Accordingly, the relationship
 between the ignition delay period and the fuel injection period can be
 restored to the target state even in case that the ignition timing is
 changed owing to low cetane number fuel or the like in a non-low
 temperature premix combustion region, thereby preventing combustion noise
 from increasing while suppressing smoke in exhaust gas.
 A fourth aspect of the present invention resides, as shown in FIG. 16, in a
 combustion control system for diesel engine, comprising a section 81 for
 detecting an ignition timing. A section 91 is provided for judging that
 the detected ignition timing is retarded relative to a target value.
 Additionally, a section 101 is provided for controlling an oxygen
 concentration in a combustion chamber so as to be decreased to a
 predetermined value when the detected ignition timing is retarded relative
 to the target value. Accordingly, a control point can be shifted into a
 region equivalent to that contains a target value even in case that it is
 difficult that the relationship between the fuel injection period and the
 ignition delay period cannot be restored to the target state in the
 non-low temperature premix combustion region, owing to an excessive
 prolongation of the ignition delay period, for example, due to low cetane
 number fuel. This effectively achieves suppressing combustion noise of the
 engine.
 A fifth aspect of the present invention resides, as shown in FIG. 17, in a
 combustion control system for a diesel engine, comprising a section 81 for
 detecting an ignition timing. A section 111 is provided for judging that
 first and second conditions are established in a predetermined engine
 operating region. The first condition is established when the detected
 ignition timing is advanced relative to a fuel injection termination
 timing. The second condition is established when it is difficult to
 control the detected ignition timing to generally coincide with or
 retarded relative to the fuel injection termination timing. Additionally,
 a section 112 is provided for controlling an amount of oxygen in a
 combustion chamber of the engine to increase when the first and second
 conditions are established. Accordingly, the oxygen amount in the
 combustion chamber can be increased in case that it is difficult to
 control the ignition timing to generally coincide with or retarded
 relative to the fuel injection termination timing under a condition where
 the actual ignition timing is advanced relative to the fuel injection
 termination timing. This can effectively reduce smoke in exhaust gas.

DETAILED DESCRIPTION OF THE INVENTION
 Referring now to FIG. 1, a first embodiment of a combustion control system
 for a diesel engine D, according to the present invention is illustrated
 by the reference character S. The diesel engine D includes an engine main
 body 1 having a plurality of cylinders (not shown) in which a plurality of
 combustion chambers (not shown) are respectively formed. The diesel engine
 D is arranged to accomplish so-called low temperature premix combustion in
 each combustion chamber under predetermined engine operating conditions.
 Such diesel engine D itself is known as disclosed in Japanese Patent
 Provisional Publication No. 8-86251.
 In such a diesel engine D, generation of NOx (nitrogen oxides) in each
 combustion chamber largely depends on combustion temperature (or a
 temperature at combustion in the combustion chamber), so that lowering the
 combustion temperature is effective for lowering an emission level of NOx.
 During the low temperature premix combustion, combustion at a low
 combustion temperature is realized by lowering an oxygen concentration in
 the combustion chamber under the action of EGR (Exhaust Gas Recirculation)
 which is accomplished by an EGR system E. The EGR system E includes an EGR
 passage 4 which connects an exhaust gas passageway 2 and an intake air
 passageway 3. A diaphragm-type EGR valve 6 is disposed in the EGR passage
 4 and arranged to control exhaust gas (EGR gas) passing through the EGR
 passage 4 in response to a vacuum controlled by a vacuum control valve 5.
 The vacuum control valve 5 is driven in response to a duty control signal
 from a control unit 41, so that a predetermined EGR rate can be obtained
 in accordance with an engine operating condition of the diesel engine D.
 The EGR rate (%) is represented by an equation of [(EGR gas quantity/fresh
 air quantity).times.100] where the EGR gas quantity is a quantity of
 exhaust gas (EGR gas) recirculated back to the intake air passageway 3
 from the exhaust gas passageway 2; and the fresh air quantity is a
 quantity of intake air passing through the intake air passage 3 to be
 supplied to the combustion chambers of the engine main body 1. In the
 diesel engine, as shown in FIG. 3, the EGR rate is, for example, set at
 100% (the maximum value) in a low engine speed and low engine load
 operating range, in which the EGR rate is lowered as the engine speed and
 the engine load increase. If a large amount of EGR gas is recirculated
 back to the engine in a high engine load operating range, the temperature
 of intake air rises thereby degrading a NOx reduction effect due to
 exhaust gas recirculation (EGR); and a so-called ignition delay period for
 injected fuel is shortened thereby, for example, making it impossible to
 realize so-called premix combustion. The ignition delay period is a period
 between a fuel injection initiation timing and an ignition timing of the
 injected fuel. In view of this, the EGR rate is lowered stepwise with the
 increased engine load and engine speed.
 A cooling device 7 for EGR gas is disposed in the EGR passage 4 and
 includes a water jacket 8 disposed around a part of the EGR passage 4. A
 part of engine coolant (cooling water) is flown through the water jacket
 8. The water jacket 8 is provided with an inlet pipe 7a through which
 engine coolant is flown into the water jacket 8 from the engine main body
 1, and an outlet pipe 7b through which engine coolant is discharged back
 to the engine main body 1. A flow control valve 9 is disposed between the
 engine main body 1 and the inlet pipe 7a of the water jacket 8 so as to
 control an amount of engine coolant circulating through the water jacket 8
 in response to a command from the control unit 41. The degree of cooling
 for EGR gas increases as the amount of engine coolant circulating through
 the water jacket 8 increases.
 A swirl valve (not shown) is movably disposed in the intake air passageway
 3 in the vicinity of an intake port for each combustion chamber. The swirl
 valve is formed with a cutout through which intake air can flow. When this
 swirl valve is closed in the low engine speed and low load operating range
 in response to a command from the control unit 41, the flow speed of
 intake air sucked into the combustion chamber is increased so that swirl
 is generated in each combustion chamber of the engine main body 1. The
 combustion chamber is of the large diameter toroidal type wherein a piston
 has a generally cylindrical piston cavity at its piston crown, though not
 shown. The piston cavity is not narrowed at its inlet or top section and
 formed at its bottom with a conical section which are coaxial with the
 piston cavity, so that no resistance is given to swirl flowing in the
 piston cavity from the outside of the piston cavity upon making its
 turning at the latter period of compression stroke of the piston while
 effectively mixing air and fuel. By virtue of the fact that the inlet
 section of the piston cavity is not narrowed, swirl generated under the
 action of the swirl valve and the like is diffused from the inside of the
 piston cavity to the outside of the piston cavity as the piston is
 descending on combustion stroke of the piston, so that swirl can be
 maintained also outside of the piston cavity.
 The engine D is provided with a fuel injection system 10 of a so-called
 common rail type. Such a common rail type fuel injection system 10 is also
 known as disclosed in "The 13th Internal Combustion Engine Symposium
 lecture papers (pages 73 to 77). The common rail type fuel injection
 system 10 will be discussed with reference to FIG. 2. The fuel injection
 system 10 includes a plurality of fuel injection nozzles 17 which are
 respectively for the cylinders (or the combustion chambers) formed in the
 engine main body 1. A supply pump 14 is provided to pressurize fuel
 supplied through a fuel supply passage 12 from a fuel tank 11. The fuel
 pressurized by the supply pump 14 is once accumulated in a pressure
 accumulator chamber (common-rail) 16 and thereafter is distributed into
 the fuel injection nozzles 17 so that a high pressure fuel is fed to each
 fuel injection nozzle 17. Each fuel injection nozzle 17 includes a needle
 valve 18 around which a nozzle chamber 19 is formed. A fuel supply passage
 20 is formed to be in communication with the nozzle chamber 19. A retainer
 21 is disposed between the needle valve 18 and a hydraulic piston 22. A
 return spring 23 is disposed around the retainer 21 to bias the needle
 valve 18 in a direction to cause the needle valve 18 to close in FIG. 2. A
 fuel supply passage 24 is formed to supply fuel onto the hydraulic piston
 22. A three-way valve (electromagnetic valve) 25 is disposed in the fuel
 supply passage 24. The fuel supply passages 20, 24 are connected to the
 pressure accumulator chamber 16.
 The three-way valve 25 is configured as follows: When the three-way valve
 25 takes its OFF position where communication is established between ports
 A and B while communication is blocked between the port B and a port C,
 the fuel supply passages 20, 24 are brought into communication with each
 other so that high pressure fuel from the pressure accumulator chamber 16
 is introduced onto the top surface of the hydraulic piston 22 and into the
 nozzle chamber 19. At this time, the needle valve 18 is in its seated
 state so that no fuel is injected through a fuel injection hole formed at
 the tip end of the fuel injection nozzle 17, because the piston has a
 pressure receiving surface larger than that of the needle valve 18. When
 the three-way valve 25 takes its ON position where communication is
 blocked between the ports A and B while communication is established
 between the ports B and C, high pressure fuel introduced onto the top
 surface of the hydraulic piston 22 is returned to the fuel tank 11 through
 a fuel return passage 28 so that a fuel pressure applied to the hydraulic
 piston 22 is lowered. As a result, the needle valve ascends so that fuel
 is injected through the fuel injection hole formed at the tip end of the
 fuel injection nozzle 17.
 When the three-way valve 25 again takes its OFF position, high pressure
 fuel from the pressure accumulator chamber 16 is introduced onto the
 hydraulic piston 22 thereby completing fuel injection from the fuel
 injection nozzle 17. In other words, the change-over timing of from the
 OFF position to the ON position of the three-way valve 25 corresponds to a
 fuel injection initiation timing at which fuel injection is initiated,
 while the time duration of the ON position of the three-way valve 25
 corresponds to a fuel injection quantity (or a quantity of fuel to be
 injected) so that the fuel injection quantity increases as the time
 duration of the ON position becomes long if the pressure in the pressure
 accumulator chamber is the same. In FIG. 2, the reference numerals 26 and
 27 designate a check valve and an orifice, respectively.
 The fuel injection system 10 further includes a pressure regulator valve 31
 disposed in a fuel return passage 13 through which fuel discharged from
 the supply pump 14 is returnable to the fuel tank 11 for the purpose of
 regulating the pressure within the pressure accumulator chamber 16. The
 pressure regulator valve 31 is adapted to allow or block flow of fuel
 through the fuel return passage 13 thereby controlling the quantity of
 fuel to be discharged to the pressure accumulator chamber 16 thus to
 regulate the pressure within the pressure regulator chamber 16. It will be
 understood that a fuel injection rate (or a quantity of fuel to be
 injected per unit time) changes in accordance with the pressure of fuel
 within the pressure accumulator chamber 16, in which the fuel injection
 rate increases as the fuel pressure within the pressure accumulator
 chamber 16 increases.
 The control unit 41 is electrically connected to the three-way valve 25 of
 the fuel injection nozzle 17 and to the pressure regulator valve 31, and
 electrically connected to an accelerator position sensor 33 for detecting
 the position of an accelerator pedal, a sensor 34 for detecting an engine
 speed of the engine and a crank angle of a crankshaft of the engine, a
 sensor 35 for detecting a cylinder to be controlled, an engine coolant
 temperature sensor 36 for detecting the temperature of the engine coolant.
 Accordingly, signals from the sensors 33, 34, 35, 36 are input to the
 control unit 41. The control unit 41 functions to calculate a target fuel
 injection quantity (or a quantity of fuel to be injected into the
 cylinder) and a target pressure within the pressure accumulator chamber 16
 in accordance with the engine speed and the accelerator pedal position,
 and makes such a feedback control through the pressure regulator valve 31
 that the pressure within the pressure accumulator chamber 16 detected by a
 pressure sensor 32 coincides with the target pressure.
 Additionally, the control unit 41 functions to so control the time duration
 of the ON position of the three-way valve 25 as to correspond to the
 calculated target fuel injection quantity, and to control the change-over
 timing of from the OFF position to the ON position of the three-way valve
 25 thereby to obtain a predetermined value of the fuel injection
 initiation timing suitable for the engine operating condition. For
 example, as shown in FIG. 4, the fuel injection initiation timing (fuel
 injection timing) is retarded to the top dead center (TDC) of the piston
 so as to increase the ignition delay period in the low engine speed and
 low engine operating range in which EGR is made at a high EGR rate. This
 retardation of the fuel injection initiation timing establishes a low
 temperature condition within the combustion chamber at a timing where
 ignition is to be made, while accomplishing a combustion whose main part
 is the premix combustion, thus suppressing generation of smoke in exhaust
 gas at a high EGR rate engine operating range. In contrast, as engine
 speed and engine load increase, the fuel injection initiation timing is
 advanced. This is because an ignition delay crank angle (obtained by
 converting the ignition delay (time) period into a crank angle) increases
 in proportion to engine speed even if the ignition delay (time) period is
 constant. Thus, the fuel injection initiation timing is advanced to obtain
 a predetermined ignition timing at a low EGR rate operating condition.
 Turning back to FIG. 1, a turbocharger T includes an exhaust gas turbine 52
 which is disposed in the exhaust gas passageway 2 downstream of a portion
 to which the EGR passage 4 is opened. A variable vane 53 is disposed at
 the scroll inlet of the turbine 52 and adapted to be driven by a step
 motor 54. It will be understood that a compressor of the turbocharger T is
 disposed in the intake air passageway 3. Here, the control unit 41
 controls the variable vane 53 through the step motor 54 in such a manner
 that the variable vane 53 takes a first vane angle (or inclined state) in
 a low engine speed operating range in order to increase the flow speed of
 exhaust gas to be introduced into the turbine 52 in a low engine speed
 operating range, and a second vane angle (or fully opened state) in a high
 engine speed operating range in order to minimize resistance of exhaust
 gas to be introduced into the turbine 52. This can provide a predetermined
 supercharged pressure (or pressure generated in the intake passage 3 by
 the turbocharger T) throughout low to high engine speed operating ranges.
 Additionally, the variable vane 53 is controlled to take a third vane
 angle for lowering the supercharged pressure in a predetermined engine
 operating condition.
 Now, the latest experiments have revealed that the most important condition
 for realizing the low temperature premix combustion is to terminate fuel
 injection during the ignition delay period (i.e., the ignition delay
 period is longer than a fuel injection period for which fuel is being
 injected), in addition to promoted dispersion of fuel under gas flow
 control. Accordingly, if the ignition delay period is changed owing to
 difference in circumferential factors such as cetane number of fuel so as
 to become shorter than the fuel injection period, the low temperature
 premix combustion cannot be accomplished and therefore exhaust gas
 purification performance inherent in the low temperature premix combustion
 cannot be obtained. For example, in case that the cetane number of fuel is
 high so that the ignition delay period is shorter than the fuel injection
 period, the low temperature premix combustion cannot be realized and
 therefore diffusive combustion is mainly made in the combustion stroke
 thereby increasing smoke in exhaust gas.
 In case that the cetane number of fuel is low, the ignition delay period is
 prolonged and therefore is preferable for the low temperature premix
 combustion. However, if the ignition timing of fuel is further retarded
 owing to use of low cetane number fuel upon having been originally
 retarded in the low temperature premix combustion, fuel consumption is
 degraded owing to lowering in combustion pressure in a high EGR rate
 operating range while causing misfire thereby increasing unburned fuel
 emission. Additionally, this increases combustion noise in a low EGR rate
 operating condition.
 In order to cope with the above problems, according to the first embodiment
 combustion control system, an actual ignition timing of fuel is detected;
 and then the detected ignition timing is controlled to coincide with or
 retarded relative to a fuel injection termination timing at which
 injection of fuel is terminated in a low temperature premix combustion
 (engine operating) region where the low temperature premix combustion is
 carried out.
 Here, discussion will be made on a control image in case that the ignition
 delay period is changed owing to cetane number of fuel, with reference to
 FIGS. 5 and 6.
 FIG. 5 shows a condition where fuel has a cetane number higher than that of
 fuel used in a matching test for obtaining control maps and the like, and
 therefore a control point shifts from a target value thereby increasing
 smoke in exhaust gas in the low temperature premix combustion region. This
 is caused for the following reasons: A difference value (fuel injection
 period--ignition delay period) on the axis of abscissas is preferable to
 be zero or minus value in order to accomplish the low temperature premix
 combustion; however, injection of whole fuel cannot be terminated during
 the ignition delay period, owing to the high cetane number of fuel (i.e.,
 the difference value takes a plus value). As a result, the rate of
 diffusive combustion increases. In addition, the fact that the oxygen
 concentration in the combustion chamber is lowered to 16% under the action
 of exhaust gas recirculation.
 In this case, in order to restore the difference value (the fuel injection
 period--the ignition delay period) to zero or a minus value (the
 difference value is restored to the target value or zero in case of FIG.
 5), the following three operations are successively carried out:
 (a) First, an injection pressure of fuel is increased to shorten the fuel
 injection period thereby minimizing the difference of the fuel injection
 period from the ignition delay period;
 (b) Secondly, the fuel injection initiation timing is delayed by an amount
 of the ignition timing advance due to the high cetane number of fuel; and
 (c) thirdly, an amount of engine coolant flowing through the EGR gas
 cooling device 7 (or through the flow control valve 9) is increased to
 lower the temperature of EGR gas, thereby prolonging the ignition delay
 period.
 Thus, by carrying out the operations (a), (b) and (c) in combination, the
 relationship between the fuel injection period and the ignition delay
 period can be restored to its target state thereby suppressing an increase
 in smoke in exhaust gas even in case that the ignition timing is advanced
 (in crank angle) relative to the fuel injection termination timing owing
 to the high cetane number of fuel in the low temperature premix combustion
 region, as shown in FIG. 5.
 It will be appreciated that it is not necessary to combine the above all
 operations (a), (b) and (c), so that only one or two of the above
 operations (a), (b) and (c) may be carried out if the relationship between
 the fuel injection period and the ignition delay period can be restored to
 the target state.
 In case of a non-low temperature premix combustion (engine operating)
 region other than the low temperature premix combustion region, the fuel
 injection initiation timing and the fuel injection pressure are controlled
 to meet target values of the ignition timing and the fuel injection
 quantity.
 Next, in case that fuel has a cetane number lower than that of fuel used in
 the matching test is used, the following two combustion regions are taken
 into consideration:
 (1) Low temperature premix combustion region
 In this region, it is preferable to prolong the ignition delay period owing
 to a low cetane number of fuel. However, there is a limit for prolonging
 the ignition delay period. If the ignition timing is excessively retarded,
 fuel consumption is degraded while increasing unburned component emission.
 Thus, in order to avoid this shortcoming, it is sufficient to advance the
 fuel injection initiation timing so as to cause the ignition timing to
 coincide with a target value.
 (2) Normal combustion region at high load (Non-low temperature premix
 combustion region)
 FIG. 6 shows a condition where the ignition delay period is prolonged owing
 to a low cetane number of fuel thereby increasing combustion noise in the
 normal combustion region. This is caused for the following reasons: The
 premix combustion is mainly accomplished under the action of prolongation
 of the ignition delay period while the oxygen concentration is higher than
 that in the low temperature premix combustion region, and therefore
 immediate combustion occurs.
 Accordingly, under such a condition, the following four operations are
 successively carried out:
 (a) First, the fuel injection initiation timing is advanced to shorten the
 ignition delay period, while lowering the fuel injection pressure to lower
 the fuel injection rate;
 (b) Secondly, the amount of engine coolant flowing through the EGR gas
 cooling device 7 (or through the flow control valve 9) is decreased to
 raise the temperature of EGR gas;
 (c) Thirdly, the EGR rate is further increased to further raise the
 temperature of intake air thereby shortening the ignition delay period;
 and
 (d) Fourthly, the variable vane of the turbocharger is controlled to
 increase the turbocharged pressure thereby to maintain the intake air
 quantity and the oxygen concentration, since an excess air factor
 (quantity of air supplied/theoretical requirement of air) and the oxygen
 concentration at target values are decreased with a rise in intake air
 temperature and an increase in EGR rate under the effects of the above (b)
 and (c).
 Thus, by carrying out the operations (a), (b), (c) and (d) in combination,
 the relationship between the fuel injection period and the ignition delay
 period can be restored to the target state thereby preventing combustion
 noise from increasing even in case that the ignition timing is changed
 owing to a low cetane number of fuel, as shown in FIG. 6.
 Also in this case, it will be appreciated that it is not necessary to
 combine all the above operations (a), (b), (c) and (d). In this
 connection, only one of the following operations (A), (B), (C) and (D) may
 be carried out if the relationship between the fuel injection period and
 the ignition delay period can be restored to the target state:
 (A) Advancing the fuel injection initiation timing (forming part of the
 above operation (a));
 (B) Lowering the fuel injection pressure (forming part of the above
 operation (a));
 (C) Combining the above operations (b) and (d); and
 (D) Combining the above operations (c) and (d).
 Next, control for attaining the effects of FIGS. 5 and 6 will be discussed
 in detail with reference to flowcharts of FIGS. 7 and 8. A control routine
 in FIGS. 7 and 8 is executed every a predetermined time.
 At a step S1, a basic fuel injection pressure P0, a basic fuel injection
 initiation timing IT0 (corresponding to the fuel injection timing in FIG.
 4), a basic EGR rate Qegr0 (corresponding to EGR rate in FIG. 3), a basic
 amount Qc0 of engine coolant flowing through the EGR gas cooling device 7,
 and a basic vane angle .theta.0 of the vane 53 of the turbocharger T are
 read. These basic values (P0, IT0, Qegr0, Qc0, .theta.0) are calculated
 upon searching maps or tables.
 At a step S2, the actual ignition timing Cst is read. Here, the actual
 ignition timing Cst is detected by a known method which is, for example,
 carried out as follows: A pressure inside the cylinder of the engine
 rapidly rises upon ignition of fuel. This pressure rise is detected by a
 pressure sensor including an piezoelectric element. The pressure sensor
 serves as an ignition timing sensor 37 as shown in FIG. 1. The ignition
 timing is detected as a timing at which a differentiated value of the
 pressure detected by the sensor 37 has reached a predetermined value or
 higher.
 At a step S3, comparison is made between the actual oxygen concentration
 O2con and a predetermined value (for example, 16%). Here, the
 predetermined value (16%) is the oxygen concentration at which the low
 temperature premix combustion is mainly made. Accordingly, it is judged
 that engine operation is in the low temperature premix combustion region
 in case that the actual oxygen concentration O2con is lower than 16%, and
 in the non-low temperature premix combustion region in case that the
 actual oxygen concentration O2con exceeds 16%. The actual oxygen
 concentration O2con can be determined by using detected values of an
 air-fuel ratio sensor (or oxygen sensor) 38 and an airflow meter 39. The
 air-fuel ratio sensor 38 and the airflow meter 39 are respectively
 disposed in the exhaust gas passageway 2 and the intake air passageway 3
 as shown in FIG. 1. It will be understood that the predetermined value
 (16%) of the oxygen concentration is different depending on engines.
 When engine operation is in the low temperature premix combustion region, a
 flow goes to a step S4 in which comparison is made between the ignition
 timing Cst and a target value of the ignition timing. Here, the ignition
 timing target value in the low temperature premix combustion region is the
 same as or retarded (in crank angle) relative to the fuel injection
 termination timing. In concrete, the ignition timing target value has been
 previously set in accordance with engine speed and engine torque (load) as
 shown in FIG. 9, so that the target value may be determined by, for
 example, searching a predetermined map in accordance with engine speed and
 engine torque.
 When the ignition timing Cst is advanced relative to the target value, the
 flow goes to steps S5, S6, S7 where calculations are made to obtain a
 correction amount .DELTA.P1 for raising the fuel injection pressure, a
 correction amount .DELTA.IT1 for retarding the fuel injection initiation
 timing, and a correction amount .DELTA.Qc1 for increasing the engine
 coolant amount flowing through the EGR gas cooling device. At a step S8,
 these correction amounts are added to the corresponding basic values P0,
 IT0, Qc0 thereby making correction to obtain the fuel injection pressure
 P, the fuel injection initiation timing IT, and the engine coolant amount
 Qc. Concerning the EGR rate Qegr and the vane angle .theta. which are
 unnecessary to be corrected, their basic values Qegr0 and the vane angle
 .theta.0 are used as they are without being corrected. At a step S9, the
 above values P, IT, Qc, Qegr and .theta. are stored in a predetermined
 address thus completing the processing of this routine.
 Concerning the calculation formula for the fuel injection initiation timing
 IT at the step S8, a standard position for fuel injection timing control
 is, for example, at a crank angle position which is considerably advanced
 relative to the most advanced value of the calculated values of the fuel
 injection initiation timing. Accordingly, the value of the fuel injection
 initiation timing is retarded relative to the above standard position, so
 that "+" in front of the .DELTA.IT1 in the calculation formula means a
 retardation in crank angle.
 When the actual ignition timing Cst is not advanced relative to the target
 value, the flow goes from the step S4 to a step S10 where comparison is
 made between the actual ignition timing Cst and a predetermined crank
 angle (for example, 15 degrees after top dead center on compression
 stroke). In order to accomplish the low temperature premix combustion, it
 is preferable to make ignition at the ignition timing retarded relative to
 the fuel injection termination timing; however, there is a limit for
 retarding the ignition timing. In this regard, if combustion is initiated
 at the ignition timing which is excessively retarded, combustion pressure
 is lowered thereby degrading fuel consumption while combustion temperature
 is lowered thereby increasing unburned component emission. Accordingly,
 the limit over which a combustion initiation timing (ignition timing) must
 not be retarded is set at 15 degrees after top dead center (on compression
 stroke) as seen from FIG. 10. Preferably, the limit is set at 15 degrees
 ATDC (after top dead center) .+-.5 degrees. While the limit has been
 described to be set at 15 degrees after top dead center on compression
 stroke in this case, it will be understood that this limit may be
 different depending upon engines.
 When the ignition timing Cst is retarded relative to 15 degrees after top
 dead center on compression stroke, the flow goes to a step S11 where a
 correction amount .DELTA.IT2 for advancing the fuel injection initiation
 timing is calculated. At a step S12, the correction amount .DELTA.IT2 is
 subtracted from the basic value IT0 thereby to obtain the fuel injection
 initiation timing IT. Concerning the fuel injection pressure P, the engine
 coolant amount Qc, the EGR rate Qegr and the vane angle .theta. which are
 unnecessary to be corrected, their basic values P0, Qc0, Qegr0 and .theta.
 are used as they are without being corrected. Thereafter, the flow goes to
 the step S9 in which the above values IT, P, Qc, Qegr and .theta. are
 stored in the predetermined address thus completing the processing of this
 routine.
 When the ignition timing Cst is not retarded relative to 15 degrees after
 top dead center on compression stroke (i.e., the actual ignition timing
 Cst coincides with the target value), the flow goes from step S10 to a
 step S13 where the basic values P, IT, Qegr, Qc and .theta. are
 respectively applied to the fuel injection pressure P, the fuel injection
 initiation timing IT, the EGR rate Qegr, the engine coolant amount Qc and
 the vane angle .theta.. Thereafter, the flow goes to the step S9 where the
 above values P, IT, Qc, Qegr and .theta. are stored in the predetermined
 address thus completing the processing of this routine.
 When engine operation is in the non-low temperature premix combustion
 region, the flow goes from the step S3 in FIG. 7 to a step S14 in FIG. 8,
 in which comparison is made between the ignition timing Cst and the target
 value. The target value of the ignition timing in the non-low temperature
 premix combustion region is not necessarily the same as that in the low
 temperature premix combustion region. For example, in case of FIG. 6, the
 ignition timing target value is positioned at a plus value in value (the
 fuel injection period--the ignition delay period), which means that the
 ignition timing is advanced relative to the fuel injection termination
 timing. This is because the ignition timing is controlled upon taking
 account of combustion noise in the non-low temperature premix combustion.
 When the ignition timing is retarded relative to the target value, the flow
 goes from the step S14 to steps 15 ,16, 17, 18, and 19 where calculations
 are made to obtain a correction amount .DELTA.IT3 for advancing the fuel
 injection initiation timing, a correction amount .DELTA.P2 for lowering
 the fuel injection pressure, and a correction amount .DELTA.Qc2 for
 decreasing the engine coolant amount flowing through the EGR gas cooling
 device, a correction amount .DELTA.Qegr for increasing the EGR rate, and a
 vane angle correction amount .DELTA..theta. for increasing the
 supercharged pressure. Thereafter, the flow goes to a step S20 where the
 correction amounts .DELTA.IT3, .DELTA.P2, .DELTA.Qc2 are respectively
 subtracted from the basic values thereby making correction to obtain the
 corrected values of the fuel injection timing IT, the fuel injection
 pressure P and the engine coolant amount Qc. Additionally, the correction
 amounts .DELTA.Qegr, .DELTA..theta. are respectively added to the basic
 values thereby making correction to obtain the EGR rate Qegr and the vane
 angle .theta.. In the calculation formula for the vane angle at the step
 S20, the supercharged pressure increases as the calculated value of the
 vane angle increases.
 FIGS. 11A to 13 illustrate a second embodiment of the combustion control
 system S according to the present invention, which is similar to the first
 embodiment combustion control system of FIGS. 1 to 10. It will be
 understood that the second embodiment combustion control system S is the
 same in structural configuration as the combustion control system shown in
 FIGS. 1 and 2. While the first embodiment has been shown and described as
 being configured to be capable of restore the relationship between the
 fuel injection period and the ignition delay period to the target state,
 the second embodiment is configured to deal with such a difficulty that
 the relationship between the fuel injection period and the ignition delay
 period cannot be restored to the target state in the non-premix combustion
 region, owing to an excessive prolongation of the ignition delay period,
 for example, due to a low cetane number of fuel.
 As discussed above, in the first embodiment, correction for restoring the
 relationship between the fuel injection period and the ignition delay
 period to the target state is accomplished as shown in FIG. 6.
 However, according to this embodiment, in case that correction of the
 ignition delay period is impossible, the following operations are carried
 out:
 (a) First, the EGR rate is increased to lower the oxygen concentration.
 This is an operation for shifting combustion into the low temperature
 premix combustion; and
 (b) The above operation (a) not only has changed the oxygen concentration
 but also has increased the temperature of intake air under the action of a
 large amount of high EGR gas thereby to shorten the ignition delay period.
 In order to avoid this shortcoming, a closing timing of each intake valve
 60 (shown in FIG. 13) of the engine E is retarded to lower an actual
 compression ratio (i.e., lowering the temperature at the terminal period
 on compression stroke) thereby prolonging the ignition delay period.
 With the above two operations (a) and (b), the control point can shift to a
 position (where combustion noise level is low) generally equal to the
 target value as shown in FIG. 11A, thus effectively suppressing combustion
 noise even in case that the correction of the ignition delay period is
 impossible. Additionally, by retarding the closing timing of each intake
 valve, intake air is discharged at the first half of compression stroke
 through the intake valve, and therefore an oxygen amount is decreased
 while the oxygen concentration is the same so that the excess air factor
 is lowered. As a result, combustion falls in the premix combustion region
 thereby preventing smoke from increasing in exhaust gas, as shown in FIG.
 11B.
 Thus, according to the second embodiment, in case that the relationship
 between the fuel injection period and the ignition delay period is
 difficult to be restored to the target state in the non-low temperature
 premix combustion owing to the excessive prolongation of the ignition
 delay period, for example, due to use of a low cetane number fuel, the
 oxygen concentration is further lowered to cause combustion to shift to
 the low temperature premix combustion while the closing timing of each
 intake valve is retarded to lower the temperature at the terminal period
 of compression stroke. This can prevent smoke in exhaust gas from
 increasing while lowering combustion noise of the engine.
 Even in case that the ignition delay period is largely prolonged owing to a
 low cetane number fuel, prolongation of the ignition delay period is
 preferable in the low temperature premix combustion region having the
 oxygen concentration of not higher than 16% as discussed in connection
 with the first embodiment, so that it is sufficient only to advance the
 fuel injection timing so as to cause the ignition timing to coincide with
 the target value.
 Control for attaining the effects of FIGS. 11A and 11B will be discussed in
 detail with reference to a flowchart of FIG. 12. A control routine in FIG.
 12 is executed every a predetermined time. The control of the flowchart in
 FIG. 12 is similar to that of the first embodiment in FIGS. 7 and 8, so
 that the same step numbers are assigned respectively to the similar steps
 to those in FIGS. 7 and 8. It will be understood that objects to be
 corrected are the fuel injection initiation timing, the EGR rate and the
 closing timing of each intake valve in the second embodiment, in which the
 values (the fuel injection pressure, the engine coolant amount and the
 vane angle) which are not necessary to be corrected are omitted.
 Additionally, in the second embodiment, only the case that the ignition
 delay period is prolonged owing to low cetane number fuel or the like is
 dealt with to be controlled, a case that the ignition delay period is
 shortened will not be discussed.
 Discussion is mainly made on parts different from the first embodiment,
 with reference to the flowchart of FIG. 12.
 When the relationship between the actual oxygen concentration O2con and the
 predetermined value of 16% is O2con&lt;16% (the non-low temperature premix
 combustion region) while the ignition timing Cst is retarded relative to
 the target value, the flow goes through the steps 3, 4 and 14 to steps 21
 and 22 where calculations are made to obtain an EGR rate correction amount
 .DELTA.Qegr2 for causing actual oxygen concentration to be not higher than
 16% and a correction amount .DELTA.IVC for retarding the closing timing of
 the intake valve. At a step S23, these correction amounts .DELTA.Qegr2,
 .DELTA.IVC are added respectively to the basic values Qegr0 and IVC0
 thereby making correction to obtain the EGR rate Qegr and the closing
 timing IVC of the intake valve. For the fuel injection initiation timing
 IT which is not necessary to be corrected, the basic value is applied as
 it is.
 Concerning the calculation formula for the closing timing of the intake
 valve, similarly to the fuel injection initiation timing, a standard
 position for intake valve closing timing control is, for example, at a
 crank angle position which is considerably advanced relative to the most
 advanced value of the calculated values of the intake valve closing
 timing. Accordingly, the value of the intake valve closing timing is
 retarded relative to the above standard position, so that "+" in front of
 the .DELTA.IVC in the calculation formula means a retardation in crank
 angle.
 In order to control the closing timing of each intake valve, a known
 variable valve timing mechanism may be used. An example of such a known
 mechanism is disclosed in Japanese Patent Provisional Publication No.
 8-254134 and shown in FIG. 13. Brief discussion will be made on the
 variable valve timing mechanism 59 with reference to FIG. 13. The engine
 main body 1 of the diesel engine E has the intake valves 60 which are
 biased in a direction to be closed under the action of valve springs 61.
 The variable valve timing mechanism 59 includes pistons 63 which are
 connected respectively to the upper ends of the intake valves 60. Each
 piston 63 defines a hydraulic chamber 62 which is to be supplied with
 hydraulic pressure. The piston 63 can descend against the bias of the
 valve spring 61 under the action of hydraulic pressure introduced into the
 hydraulic chamber 62.
 Hydraulic fluid discharged from an oil pump 64 is selectively supplied from
 an accumulator 65 into oil passages 68, 69 through inlet-side
 electromagnetic change-over valves 66, 67. The hydraulic fluid supplied to
 the oil passage 68, 69 is selectively supplied to the hydraulic chambers
 62, 62 for Nos. 1 and 4 cylinders of the engine through a rotary valve 70,
 and to the hydraulic chambers 62, 62 for Nos. 2 and 3 cylinders of the
 engine through a rotary valve 71. Consequently, the intake valve 60 for
 the No. 1 cylinder, the intake valve 60 for the No. 4 cylinder, the intake
 valve 60 for the No. 2 cylinder and the intake valve 60 for the No. 3
 cylinder are successively opened. Each rotary valve 70, 71 is rotatable in
 timed relation to engine speed of the engine.
 The hydraulic fluid in each hydraulic chamber 62 is selectively discharged
 from the oil passage 68, 69 through one of outlet-side electromagnetic
 change-over valves 73, 74 to a tank 75, so that the intake valves 60 are
 successively closed. The closing timing of each intake valve 60 is
 variably regulated by controlling the outlet-side electromagnetic
 change-over valves 73, 74. Accordingly, in case of using the variable
 valve timing mechanism 59 (shown in FIG. 13) in the combustion control
 system S of the present invention, the outlet-side electromagnetic
 change-over valves 73, 74 are controlled in accordance with the values of
 the intake valve closing timing IVC obtained upon execution of the flow in
 FIG. 12.
 Experiments have revealed that there exists a case in which it is difficult
 to control the ignition timing Cst to coincide with or retarded relative
 to the fuel injection termination timing under a condition where the
 actual ignition timing Cst is advanced relative to the fuel injection
 termination timing in the low temperature premix combustion, though not
 discussed. In such a case, a control for increasing the oxygen amount in
 the combustion chamber is made (for example, by increasing the
 supercharged pressure by the turbocharger or by decreasing the EGR rate),
 giving up carrying out the low temperature premix combustion. This shits
 combustion into normal combustion whose major part is diffusive
 combustion, thereby decreasing smoke in exhaust gas.
 While the embodiments of the present invention have been shown and
 described as being arranged such that the low temperature premix
 combustion is carried out, it will be understood that the principle of the
 present invention may be applied to diesel engines in which no low
 temperature premix combustion is made.
 Although the oxygen amount or concentration in the combustion chamber has
 been shown and described as being changed by the EGR rate in the
 combustion control systems of the embodiments, it will be appreciated that
 the principle of the present invention may be applied to other combustion
 control systems for diesel engines, such as those configured such that the
 oxygen amount is changed, for example, by using an oxygen permeable
 membrane.