Abstract:
A method for controlling a multi-cylinder direct-injection internal combustion engine includes establishing an external exhaust gas recirculation valve from a present position to a target position to achieve a target exhaust gas recirculation, and synchronizing combustion initiation timing with actual exhaust gas recirculation as actual exhaust gas recirculation changes from a first exhaust gas recirculation value corresponding to the present position of the external exhaust gas recirculation valve to a second exhaust gas recirculation value corresponding to the target position of the external exhaust gas recirculation valve.

Description:
TECHNICAL FIELD 
     This disclosure relates to operation and control of homogeneous-charge compression-ignition (HCCI) engines. 
     BACKGROUND 
     The statements in this section merely provide background information related to the present disclosure and may not constitute prior art. 
     Internal combustion engines, especially automotive internal combustion engines, generally fall into one of two categories: spark ignition and compression ignition. Spark ignition engines, such as gasoline engines, introduce a fuel/air mixture into the combustion cylinders, which is then compressed in the compression stroke and ignited by a spark plug. Compression ignition engines, such as diesel engines, introduce or inject pressurized fuel into a combustion cylinder near top dead center (TDC) of the compression stroke, which ignites upon injection. Combustion for both gasoline engines and diesel engines involves premixed or diffusion flames that are controlled by fluid mechanics. Each type of engine has advantages and disadvantages. In general, gasoline engines produce fewer emissions but are less efficient. In general, diesel engines are more efficient but produce more emissions. 
     More recently, other types of combustion methodologies have been introduced for internal combustion engines. One of these combustion concepts is known in the art as the homogeneous charge compression ignition (HCCI). The HCCI combustion mode includes a distributed, flameless, auto-ignition combustion process that is controlled by oxidation chemistry, rather than by fluid mechanics. In a typical engine operating in HCCI combustion mode, the cylinder charge is nearly homogeneous in composition and temperature at intake valve closing time. The typical engine operating in the HCCI combustion mode can further operate using stratified charge fuel injection to control and modify the combustion process, including using stratified charge combustion to trigger the HCCI combustion. Because auto-ignition is a distributed kinetically-controlled combustion process, the engine operates at a very dilute fuel/air mixture (i.e., lean of a fuel/air stoichiometric point) and has a relatively low peak combustion temperature, thus forming extremely low nitrous oxides (NOx) emissions. The fuel/air mixture for auto-ignition is relatively homogeneous, as compared to the stratified fuel/air combustion mixtures used in diesel engines. Therefore, the rich zones that form smoke and particulate emissions in diesel engines are substantially eliminated. Because of this very dilute fuel/air mixture, an engine operating in the auto-ignition combustion mode can operate unthrottled to achieve diesel-like fuel economy. The HCCI engine can operate at stoichiometry with substantial amounts of exhaust gas recirculation (EGR) to achieve effective combustion. 
     There is no direct control of start of combustion for an engine operating in the auto-ignition mode, as the chemical kinetics of the cylinder charge determine the start and course of the combustion. Chemical kinetics are sensitive to temperature and, as such, the controlled auto-ignition combustion process is sensitive to temperature. An important variable affecting the combustion initiation and progress is the effective temperature of the cylinder structure, i.e., temperature of cylinder walls, head, valve, and piston crown. Additionally, spark-assisted ignition is known to facilitate combustion in certain operating ranges. 
     Operation within an HCCI mode at higher loads can be challenging, as energy present within the combustion chamber increases with increasing load. This increasing energy, exhibited for example by higher temperatures within the air fuel charge being combusted, increases likelihood of the air fuel charge combusting before the desired combustion point, resulting in an undesirable pressure wave or ringing from the combustion chamber. 
     SUMMARY 
     A method for controlling a multi-cylinder direct-injection internal combustion engine includes establishing an external exhaust gas recirculation valve from a present position to a target position to achieve a target exhaust gas recirculation, and synchronizing combustion initiation timing with actual exhaust gas recirculation as actual exhaust gas recirculation changes from a first exhaust gas recirculation value corresponding to the present position of the external exhaust gas recirculation valve to a second exhaust gas recirculation value corresponding to the target position of the external exhaust gas recirculation valve. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       One or more embodiments will now be described, by way of example, with reference to the accompanying drawings, in which: 
         FIG. 1  is a schematic drawing of an exemplary engine system, in accordance with the present disclosure; 
         FIG. 2  is a graph depicting a heat release rate profile during spark-assisted HCCI combustion, in accordance with the present disclosure; 
         FIG. 3  is a plot of fuel injection mass  302  versus external exhaust gas recirculation (EGR) percentage  301  required to maintain desired stoichiometry for spark-assisted HCCI combustion, in accordance with the present disclosure; 
         FIG. 4  schematically illustrates an external EGR compensation controller  400  for adjusting fuel injection and/or spark timing to be synchronized with emptying and filling of external EGR during load transients in spark-assisted HCCI combustion mode, in accordance with the present disclosure; 
         FIG. 5  graphically depicts experimental and derived data from an exemplary engine, depicting fuel injected mass  501 , fuel injection timing  503 , spark ignition timing  505 , EGR valve opening  507 , CA50 (i.e., crank angle location of 50% fuel mass burn)  509  and average IMEP  511  versus time  502 , in accordance with the present disclosure; and 
         FIG. 6  graphically depicts experimental and derived data from an exemplary engine, depicting fuel injected mass  601 , fuel injection timing  603 , spark ignition timing  605 , EGR valve opening  607 , CA50 (i.e., crank angle location of 50% fuel mass burn)  609  and average IMEP  611  versus time  602 , in accordance with the present disclosure. 
     
    
    
     DETAILED DESCRIPTION 
     Referring now to the drawings, wherein the depictions are for the purpose of illustrating certain exemplary embodiments only and not for the purpose of limiting the same,  FIG. 1  schematically shows an exemplary internal combustion engine  10  and an accompanying control module  5  that have been constructed in accordance with an embodiment of the disclosure. The engine  10  is selectively operative in a plurality of combustion modes, including a controlled auto-ignition (HCCI) combustion mode and a homogeneous spark-ignition (SI) combustion mode. The engine  10  is selectively operative at a stoichiometric air/fuel ratio and at an air/fuel ratio that is primarily lean of stoichiometry. It is appreciated that the concepts in the disclosure can be applied to other internal combustion engine systems and combustion cycles. 
     In one embodiment the engine  10  can be coupled to a transmission device to transmit tractive power to a driveline of a vehicle. The transmission can include a hybrid transmission including torque machines operative to transfer tractive power to a driveline. 
     The exemplary engine  10  includes a multi-cylinder direct-injection four-stroke internal combustion engine having reciprocating pistons  14  slidably movable in cylinders  15  which define variable volume combustion chambers  16 . Each piston  14  is connected to a rotating crankshaft  12  by which linear reciprocating motion is translated to rotational motion. An air intake system provides intake air to an intake manifold  29  which directs and distributes air into intake runners of the combustion chambers  16 . The air intake system includes airflow ductwork and devices for monitoring and controlling the airflow. The air intake devices preferably include a mass airflow sensor  32  for monitoring mass airflow and intake air temperature. A throttle valve  34  preferably includes an electronically controlled device that is used to control airflow to the engine  10  in response to a control signal (ETC) from the control module  5 . A pressure sensor  36  in the intake manifold  29  is configured to monitor manifold absolute pressure and barometric pressure. An external flow passage recirculates exhaust gases from engine exhaust to the intake manifold  29 , having a flow control valve referred to as an exhaust gas recirculation (EGR) valve  38 . The control module  5  is operative to control mass flow of exhaust gas to the intake manifold  29  by controlling opening of the EGR valve  38 . 
     Airflow from the intake manifold  29  into the combustion chamber  16  is controlled by one or more intake valve(s)  20 . Exhaust flow out of the combustion chamber  16  is controlled by one or more exhaust valve(s)  18  to an exhaust manifold  39 . The engine  10  is equipped with systems to control and adjust openings and closings of the intake and exhaust valves  20  and  18 . In one embodiment, the openings and closings of the intake and exhaust valves  20  and  18  can be controlled and adjusted by controlling intake and exhaust variable cam phasing/variable lift control (VCP/VLC) devices  22  and  24  respectively. The intake and exhaust VCP/VLC devices  22  and  24  are configured to control and operate an intake camshaft  21  and an exhaust camshaft  23 , respectively. The rotations of the intake and exhaust camshafts  21  and  23  are linked to and indexed to rotation of the crankshaft  12 , thus linking openings and closings of the intake and exhaust valves  20  and  18  to positions of the crankshaft  12  and the pistons  14 . 
     The intake VCP/VLC device  22  preferably includes a mechanism operative to switch and control valve lift of the intake valve(s)  20  and variably adjust and control phasing of the intake camshaft  21  for each cylinder  15  in response to a control signal (INTAKE) from the control module  5 . The exhaust VCP/VLC device  24  preferably includes a controllable mechanism operative to variably switch and control valve lift of the exhaust valve(s)  18  and variably adjust and control phasing of the exhaust camshaft  23  for each cylinder  15  in response to a control signal (EXHAUST) from the control module  5 . 
     The intake and exhaust VCP/VLC devices  22  and  24  each preferably includes a controllable two-step variable lift control (VLC) mechanism operative to control magnitude of valve lift, or opening, of the intake and exhaust valve(s)  20  and  18 , respectively, to one of two discrete steps. The two discrete steps preferably include a low-lift valve open position (about 4-6 mm in one embodiment) preferably for load speed, low load operation, and a high-lift valve open position (about 8-13 mm in one embodiment) preferably for high speed and high load operation. The intake and exhaust VCP/VLC devices  22  and  24  each preferably includes a variable cam phasing (VCP) mechanism to control and adjust phasing (i.e., relative timing) of opening and closing of the intake valve(s)  20  and the exhaust valve(s)  18  respectively. Adjusting the phasing refers to shifting opening times of the intake and exhaust valve(s)  20  and  18  relative to positions of the crankshaft  12  and the piston  14  in the respective cylinder  15 . The VCP mechanisms of the intake and exhaust VCP/VLC devices  22  and  24  each preferably has a range of phasing authority of about 60°-90° of crank rotation, thus permitting the control module  5  to advance or retard opening and closing of one of intake and exhaust valve(s)  20  and  18  relative to position of the piston  14  for each cylinder  15 . The range of phasing authority is defined and limited by the intake and exhaust VCP/VLC devices  22  and  24 . The intake and exhaust VCP/VLC devices  22  and  24  include camshaft position sensors to determine rotational positions of the intake and the exhaust camshafts  21  and  23 . The VCP/VLC devices  22  and  24  are actuated using one of electro-hydraulic, hydraulic, and electric control force, controlled by the control module  5 . 
     The engine  10  includes a fuel injection system, including a plurality of high-pressure fuel injectors  28  each configured to directly inject a mass of fuel into one of the combustion chambers  16  in response to a signal from the control module  5 . The fuel injectors  28  are supplied pressurized fuel from a fuel distribution system. 
     The engine  10  includes a spark-ignition system by which spark energy can be provided to a spark plug  26  for igniting or assisting in igniting cylinder charges in each of the combustion chambers  16  in response to a signal (IGN) from the control module  5 . 
     The engine  10  is equipped with various sensing devices for monitoring engine operation, including a crank sensor  42  having output RPM and operative to monitor crankshaft rotational position, i.e., crank angle and speed, in one embodiment a combustion sensor  30  configured to monitor combustion, and an exhaust gas sensor  40  configured to monitor exhaust gases, e.g. an air/fuel ratio sensor. The combustion sensor  30  includes a sensor device operative to monitor a state of a combustion parameter and is depicted as a cylinder pressure sensor operative to monitor in-cylinder combustion pressure. The output of the combustion sensor  30  and the crank sensor  42  are monitored by the control module  5  which determines combustion phasing, i.e., timing of combustion pressure relative to the crank angle of the crankshaft  12  for each cylinder  15  for each combustion cycle. The combustion sensor  30  can also be monitored by the control module  5  to determine a mean-effective-pressure (IMEP) for each cylinder  15  for each combustion cycle. Preferably, the engine  10  and control module  5  are mechanized to monitor and determine states of IMEP for each of the engine cylinders  15  during each cylinder firing event. Alternatively, other sensing systems can be used to monitor states of other combustion parameters within the scope of the disclosure, e.g., ion-sense ignition systems, and non-intrusive cylinder pressure sensors. 
     Control module, module, controller, control unit, processor and similar terms mean any suitable one or various combinations of one or more of Application Specific Integrated Circuit(s) (ASIC), electronic circuit(s), central processing unit(s) (preferably microprocessor(s)) and associated memory and storage (read only, programmable read only, random access, hard drive, etc.) executing one or more software or firmware programs, combinational logic circuit(s), input/output circuit(s) and devices, appropriate signal conditioning and buffer circuitry, and other suitable components to provide the described functionality. The control module  5  has a set of control algorithms, including resident software program instructions and calibrations stored in memory and executed to provide the desired functions. The algorithms are preferably executed during preset loop cycles. Algorithms are executed, such as by a central processing unit, and are operable to monitor inputs from sensing devices and other networked control modules, and execute control and diagnostic routines to control operation of actuators. Loop cycles may be executed at regular intervals, for example each 3.125, 6.25, 12.5, 25 and 100 milliseconds during ongoing engine and vehicle operation. Alternatively, algorithms may be executed in response to occurrence of an event. 
     In operation, the control module  5  monitors inputs from the aforementioned sensors to determine states of engine parameters. The control module  5  is configured to receive input signals from an operator (e.g., via an accelerator pedal and a brake pedal,) to determine a torque request (To_req). It will be appreciated that the torque request can be in response to an operator input (e.g., via the accelerator pedal and the brake pedal) or the torque request can be in response to an auto start condition monitored by the control module  5 . The control module  5  monitors the sensors indicating the engine speed and intake air temperature, and coolant temperature and other ambient conditions. 
     The control module  5  executes algorithmic code stored therein to control the aforementioned actuators to form the cylinder charge, including controlling throttle position, spark-ignition timing, fuel injection mass and timing, EGR valve position opening to control flow of recirculated exhaust gases, and intake and/or exhaust valve timing and phasing on engines so equipped. Valve timing and phasing can include negative valve overlap (NVO) and lift of exhaust valve reopening (in an exhaust re-breathing strategy) in one embodiment. The control module  5  can operate to turn the engine  10  on and off during ongoing vehicle operation, and can operate to selectively deactivate a portion of the combustion chambers  16  or a portion of the intake and exhaust valves  20  and  18  through control of fuel and spark and valve deactivation. The control module  5  can control air/fuel ratio based upon feedback from the exhaust gas sensor  40 . 
     During engine operation, the throttle valve  34  is preferably substantially wide-open in the controlled auto-ignition (HCCI) combustion modes, e.g., single and double injection controlled auto-ignition (HCCI) combustion modes, with the engine  10  controlled at a lean air/fuel ratio. Substantially wide-open throttle can include operating fully un-throttled, or slightly throttled to create a vacuum in the intake manifold  29  to affect EGR flow. In one embodiment, in-cylinder EGR mass is controlled to a high dilution rate. The intake and exhaust valves  20  and  18  are in the low-lift valve position and the intake and exhaust lift timing operate with NVO. One or more fuel injection events can be executed during an engine cycle including at least one fuel injection event during a compression phase. 
     During engine operation in the homogeneous spark-ignition (SI) combustion mode, the throttle valve  34  is controlled to regulate the air flow. The engine  10  is controlled to a stoichiometric air/fuel ratio, and the intake and exhaust valves  20  and  18  are in the high-lift valve open position and the intake and exhaust lift timing operate with a positive valve overlap. Preferably, a fuel injection event is executed during compression phase of an engine cycle, preferably substantially before TDC. Spark ignition is preferably discharged at a predetermined time subsequent to the fuel injection when air charge within the cylinder is substantially homogeneous. 
     Referring to  FIG. 2 , a graph depicting a heat release rate profile  200  during spark-assisted HCCI (auto-ignition) combustion is illustrated in accordance with the present disclosure. The x-axis  202  represents a Mass Fraction Burn (%) and the y-axis  201  represents the Heat Release Rate (J/deg). The high-load operating limit of controlled auto-ignition (HCCI) can be extended by spark igniting the air fuel charge prior to a auto-ignition point designated by dashed line  60 . Igniting the air fuel charge initiates flame propagation, wherein flame propagation is utilized to extend the high-load operating limit by retarding auto-ignition and thus achieving acceptable combustion noise. During high-load operation, the spark-assisted auto-ignition includes delivering the fuel mass to the engine using a single injection during an intake stroke, spark-igniting the injected fuel mass during a compression stroke, initiating flame propagation and auto-igniting the remainder of the injected fuel mass when the temperature of the cylinder charge increases by the flame propagation to a temperature sufficient for auto-ignition. Flame propagation occurs up to the dashed line  60  and is denoted by arrow  58 . Auto-ignition occurs at the dashed line  60  and is denoted by arrow  62 . As will become apparent, combustion noise and combustion misfires can be substantially reduced during engine load transients while maintaining robust combustion stability by increasing or decreasing the spark timing and/or the injected fuel mass timing (i.e., combustion initiation timing). Specifically, retarding the spark timing and/or the injected fuel mass timing retards combustion phasing and reduces combustion noise. Advancing the spark timing and/or the injected fuel mass timing advances combustion phasing and reduces combustion misfire and partial combustion burn. It is appreciated that the spark-ignited air fuel charge can include an air fuel ratio substantially at stoichiometric to utilize a three-way catalytic converter to meet desirable NOx emission standards. 
     Embodiments discussed herein utilize a control strategy for maintaining robust combustion during load transients while operating the engine in a controlled auto-ignition (HCCI) mode including spark-assisted ignition during high-load operation. It will be appreciated that the control strategy is not limited to high-load auto-ignition (HCCI) operation and can be similarly applied to low- and medium-load auto-ignition (HCCI) operation with or without spark-assisted ignition. 
     Load transients can occur in response to an operator torque request and can include rapid load transients in response to a rapid operator torque request. It will be understood that engine load corresponds to injected fuel mass required to achieve the operator torque request. Engine load and injected fuel mass will be used interchangeably herein. The injected fuel mass is increased or decreased to meet an increased or decreased operator torque request, respectively. As aforementioned, operator torque requests are monitored by the control module  5  and can include accelerator pedal and brake pedal inputs. When an engine load transient occurs, the amount of external EGR must be adjusted to achieve a value corresponding to the engine load (i.e., injected fuel mass) to maintain a desired air fuel ratio and acceptable combustion noise. For instance, the desired air fuel ratio can be substantially stoichiometric to meet desired NOx emission levels utilizing a three-way catalytic converter at high engine load. The injected fuel mass corresponding to the engine load transient is a mass sufficient to achieve the operator torque request. The amount of external EGR is adjusted utilizing the control module  5  to control mass flow of exhaust gas to the intake manifold  29  by controlling the opening percentage of the EGR valve  38 . Therefore, in response to an injected fuel mass transient (i.e., engine load transient) associated with a transient operator torque request, the external EGR valve  38  percentage opening is adjusted to achieve a value corresponding to the injected fuel mass transient to maintain the desired air fuel ratio (e.g., substantially stoichiometric) and acceptable combustion noise. The amount of external EGR will further be referred to as the external EGR percentage. 
     Referring now to  FIG. 3 , a plot of injected fuel mass versus external EGR percentage required to maintain a desired air fuel ratio (e.g., stoichiometric) and acceptable combustion noise while achieving the best fuel efficiency and emissions for spark-assisted HCCI combustion is illustrated in accordance with the present disclosure. The x-axis  302  represents Fuel Mass (mg/cycle) and the y-axis  301  represents external EGR percentage (%). Line  300  represents the desired external EGR percentage. As demonstrated for fuel exceeding at a specified mass, the external EGR percentage decreases as the fuel mass increases. 
     The external EGR has relatively slow dynamics (external EGR dynamics) compared to fuel mass injections. In other words, it takes longer to adjust the external EGR percentage than it does to adjust the injected fuel mass. The external EGR dynamics are associated with a rate for the adjusted external EGR percentage to achieve a value corresponding to the injected fuel mass. A difference in the amount of external EGR percentage required for two different injected fuel masses during engine load transients can have a significant impact on the combustion stability during such transients. For instance, if the injected fuel mass was rapidly increased due to an increased operator torque request, the engine would be initially operated with excessive amount of external EGR percentage due to slow emptying dynamics when the external EGR percentage is decreasing until the emptying dynamics achieve the external EGR percentage value corresponding to the increased injected fuel mass. The slow emptying dynamics can retard combustion phasing more than desired resulting in partial combustion burn or combustion misfire. Similarly, if the injected fuel mass was rapidly decreased due to a decreased operator torque request, the engine would be initially operated with an insufficient amount of external EGR percentage due to slow filling dynamics when the external EGR percentage is increasing until the filling dynamics achieve an external EGR percentage value corresponding to the decreased injected fuel mass to maintain the desired air fuel ratio. The slow filling dynamics can advance combustion phasing more than desired resulting in excessive combustion noise. 
     To compensate for the slow external EGR dynamics, combustion initiation timing is adjusted at a rate according to the external EGR dynamics (e.g., emptying or filling dynamics) to maintain robust combustion and substantially prevent combustion misfires and excessive combustion noise. In other words, the combustion initiation timing is synchronized with the external EGR dynamics (e.g., emptying or filling dynamics) to maintain robust combustion and prevent combustion misfires and excess combustion noise rather than synchronizing the combustion initiation timing with the injected fuel mass transient (i.e., load transient) based on static calibration because the external EGR percentage cannot be adjusted quickly due to slow external EGR dynamics. The combustion initiation timing can include injection timing and spark timing. For instance, if the injected fuel mass was rapidly increased due to an operator torque request and the external EGR percentage is decreased to achieve a value corresponding to the increased injected fuel mass, combustion initiation timing can be decreased at a slow rate to compensate for slow emptying dynamics. Slowly decreasing the combustion initiation timing includes slowly retarding combustion phasing to substantially reduce combustion misfire and combustion partial burns resulting from the initially excessive external EGR percentage. Similarly, if the injected fuel mass was rapidly decreased due to a decreased operator torque request and the external EGR percentage is increased to achieve a value corresponding to the decreased injected fuel mass, combustion initiation timing can be increased at a slow rate to compensate for slow filling dynamics. The slowly increasing combustion initiation timing includes slowly advancing combustion phasing to substantially prevent excessive combustion noise resulting from the initially insufficient external EGR percentage. 
     As aforementioned, combustion initiation timing can include fuel injection timing and spark timing. It is appreciated that adjusting fuel injection timing and spark timing can compensate for the effects that slow external EGR dynamics have on combustion phasing. Combustion phasing describes the progression of combustion in a cycle as measured by the crank angle of the cycle. One metric to judge combustion phasing is CA50 or the crank angle at which 50% of the air fuel charge is combusted. Properties of a combustion cycle, such as efficiency, are affected by CA50 of the cycle. However, other factors such as injection and spark timing can affect CA50. Injection and spark timing can both be adjusted and synchronized with the external EGR dynamics to compensate for the slow external EGR dynamics, or either one of injection timing or spark timing can be adjusted and synchronized with the external EGR dynamics to compensate for slow external EGR dynamics. 
     Adjusting injection timing can be utilized to modulate resulting combustion phasing from slow external EGR dynamics. The effect of injection timing upon combustion phasing depends upon the resulting conditions within the combustion chamber. For example, later or decreased injection timing can cause combustion to start later, thereby retarding combustion phasing. Therefore, when engine load is increased associated with an increased operator torque request and the external EGR percentage is decreased to achieve the value corresponding to the increased engine load, decreasing the fuel injection timing at a rate synchronized with the emptying dynamics of the external EGR percentage can compensate for an initially excessive external EGR percentage resulting from slow emptying dynamics until the adjusted external EGR percentage achieves the value corresponding to the increased engine load. In another example, increased or earlier injection can cause combustion to start earlier, thereby advancing combustion phasing. Therefore, when engine load is decreased associated with a decreased operator torque request and the external EGR percentage is increased to achieve the value corresponding to the decreased engine load, increasing the fuel injection timing at a rate synchronized with the filling dynamics of the external EGR percentage can compensate for an initially insufficient external EGR percentage resulting from slow filling dynamics until the adjusted external EGR percentage achieves the value corresponding to the decreased engine load. 
     Spark timing can also be utilized to modulate resulting combustion phasing resulting from slow external EGR dynamics. Spark-assisted ignition of auto-ignition (HCCI) combustion includes utilizing a spark to create combustion within the combustion chamber of an air fuel charge not yet at an energy level conducive to controlled auto-ignition (HCCI). The spark induced combustion creates a release of energy within the combustion chamber including a pressure wave. This energy release propagates to the remainder of the combustion chamber and facilitates the remainder of the air fuel charge to achieve auto-ignition (HCCI). While auto-ignition most ideally operates without spark ignition, circumstances are known wherein spark-assisted auto-ignition operation is desirable. For example, in cold start or low speed and low load conditions, spark-assisted HCCI utilizes the energy release from the spark ignition to facilitate auto-ignition of the charge in a region wherein auto-ignition might be unstable or not possible. In the present circumstances, to enable HCCI operation at higher loads, spark-assisted ignition can be used to begin combustion of the charge to initiate controlled auto-ignition (HCCI), thereby allowing control of combustion phasing through modulation of the spark timing. Testing has shown that, when such selection is possible according to injection timing, resulting CA50, and other related parameters, selection of advanced spark timing can facilitate combustion of 20% of the air fuel charge in advance of the initiation of auto-ignition. In an example, when the engine load is increased associated with an increased torque request and the external exhaust gas recirculation percentage is decreased to achieve the value corresponding to the increased load, spark timing can be selected to combust a portion of an air fuel charge to initiate auto-ignition timing. The spark timing is decreased at a rate synchronized with the emptying dynamics of the external EGR percentage to compensate for an initially excessive external EGR percentage resulting from slow emptying dynamics until the adjusted external EGR percentage achieves the value corresponding to the increased load. In another example, when the engine load is decreased corresponding to a decreased torque request and the external exhaust gas recirculation percentage is increased to achieve the value corresponding to the decreased engine load, spark timing can be selected to combust a portion of an air fuel charge to initiate auto-ignition timing. The spark timing is increased at a rate synchronized with the filling dynamics to compensate for an initially insufficient external EGR percentage resulting from slow filling dynamics until the adjusted external EGR percentage achieves the value corresponding to the decreased engine load. 
     Referring now to  FIG. 4 , a compensation controller  400  for adjusting combustion initiation timing to be synchronized with external EGR dynamics (e.g., filling/emptying dynamics) during engine load transients in spark-assisted HCCI (auto-ignition) combustion mode is illustrated in accordance with the present disclosure. The compensation controller includes a combustion mode determination module  402 , a calibration table module  404 , a time constant determination module  406 , a low pass filter module  408  and a combustion initiation timing switch module  410 . A desired load transient  412  and associated desired engine speed is input to the combustion mode determination module  402 , the calibration table module  404  and the time constant determination module  406 . The desired load transient is based upon, for example, the operator torque request, wherein the operator torque request can include operator inputs to actuators including an accelerator pedal and a brake pedal, as mentioned above. It is further appreciated that the compensation controller  400  is associated with the control module  5 . If the combustion mode determination module  402  determines spark-assisted auto-ignition mode based upon the desired load transient  412  and associated engine speed, a signal  414  indicating spark-assisted auto-ignition mode is input to the switch module  410  and the low-pass filter module  408 . The calibration table  404  inputs a signal  416  indicating static calibration of the combustion initiation timing into the low-pass filter module  408  and the switch module  410 , wherein the signal  416  is based on the desired load transient  412  and associated engine speed. The time constant determination module  406  generates a time constant signal  418  based on the desired load transient  412  and associated engine speed, wherein the time constant signal  418  is input to the low-pass filter module  408 . The low pass filter module  408  generates a filtered signal  420  that is input to the switch module  410 . The filtered signal  420  is based on the time constant signal  418  and the signal  416  indicating static calibration of the combustion initiation timing. The switching module  410  determines the adjusted combustion initiation timing  422  synchronized with external EGR dynamics to thereby compensate for slow external EGR dynamics during load transients. The adjusted combustion initiation timing  422  includes fuel injection timing and spark timing and is based upon the signal  414  indicating spark-assisted auto-ignition mode, the signal  416  indicating static calibration of the combustion initiation timing and the time constant signal  418 . 
     Referring now to  FIG. 5 , experimental and derived data from an exemplary engine are illustrated, depicting injected fuel mass  501 , fuel injection timing  503  (before TDC), spark ignition timing  505  (before TDC), EGR valve opening  507 , CA50 (i.e., crank angle location of 50% fuel mass burn)  509 , and average IMEP  511  versus time  502  in accordance with the present disclosure. In each of the data plots  503 ,  505 ,  507 ,  509  and  511 , the dashed profile lines include adjustments to compensate for slow external EGR dynamics, whereas the solid profile lines do not include adjustments to compensate for slow external EGR dynamics. When an injected fuel mass is increased  500 , compensated fuel injection timing  504  is plotted against fuel injection timing  506  without compensation; compensated spark ignition timing  508  is plotted against spark ignition timing  510  without compensation; EGR valve opening  512  with compensation is plotted against EGR valve opening  514  without compensation which is adjusted in real-time to maintain a desired CA 50 profile  516 ; and compensated IMEP  522  is plotted against IMEP  524  without compensation. It is appreciated that compensated CA 50 profile  520  maintains the desired CA 50 profile  516  when compensation for slow external EGR emptying dynamics is utilized, as opposed to CA 50 profile  518  without compensating for the external EGR emptying dynamics. 
     Referring now to  FIG. 6 , experimental and derived data from an exemplary engine are illustrated, depicting injected fuel mass  601 , fuel injection timing  603  (before TDC), spark ignition timing  605  (before TDC), EGR valve opening  607 , CA50 (i.e., crank angle location of 50% fuel mass burn)  609 , and average IMEP  611  versus time  602  in accordance with the present disclosure. In each of the data plots  603 ,  605 ,  607 ,  609  and  611 , the dashed profile lines include adjustments to compensate for slow external EGR dynamics, whereas the solid profile lines do not include adjustments to compensate for slow external EGR dynamics. When an injected fuel mass is decreased  600 , compensated fuel injection timing  604  is plotted against fuel injection timing  606  without compensation; compensated spark ignition timing  608  is plotted against spark ignition timing  610  without compensation; EGR valve opening  612  with compensation is plotted against EGR valve opening  614  without compensation which is adjusted in real-time to maintain a desired CA 50 profile  616 ; and compensated IMEP  622  is plotted against IMEP  624  without compensation. It is appreciated that compensated CA 50 profile  618  more closely maintains the desired CA 50 profile  616  when compensation for slow external EGR emptying dynamics is utilized, as opposed to CA 50 profile  620  without compensating for the external EGR emptying dynamics. 
     The disclosure has described certain preferred embodiments and modifications thereto. Further modifications and alterations may occur to others upon reading and understanding the specification. Therefore, it is intended that the disclosure not be limited to the particular embodiment(s) disclosed as the best mode contemplated for carrying out this disclosure, but that the disclosure will include all embodiments falling within the scope of the appended claims.