Abstract:
A method of controlling an intake air passage of an internal combustion engine is provided. The intake air passage cyclically communicates to a combustion chamber of the internal combustion engine, thereby inducting fresh air into said combustion chamber. The cyclic communication of the intake air passage to the combustion chamber generates a pressure wave in the intake air passage. The method comprises reducing an effective length of a transmission path of the pressure wave in an upstream direction of the intake air passage as a desired air flow to the combustion chamber decreases. In accordance with the method, the effective length of the pressure wave transmission path is reduced as desired air flow decreases. With the reduced effective length, the pressure wave bounces back and forth between ends of the transmission path more often before the next cyclic communication. The more bouncing attenuates the pressure wave at the next cyclic communication. Therefore, the cylinder air charge can be stabilized when the desired air flow is decreased.

Description:
BACKGROUND 
   The present description relates to an intake system of an internal combustion engine, and more particularly to control of a variable intake system and a variable valve lift system. 
   There is known and presented, for example in Japanese Patent Application Publication JP2004-137982A, a variable intake system of an internal combustion engine, which is capable of adjusting an effective length of an intake air passage to be more suitable for a current engine speed. It utilizes a pulse wave that a motion of air in the intake passage just upstream of an intake valve generates. Conventionally, an intake valve begins to open at a beginning of an intake stroke of a cylinder cycle. When a piston in an engine cylinder passes its top dead center and then descends, the intake valve is substantially or effectively open, and air is sucked into the cylinder. Then, the piston descends, and a negative pressure is generated in an intake port throat which is just upstream of the intake valve. 
   The negative pressure generates a pressure wave. It transmits upstream in the intake passage. Then, the pressure wave reflects at a larger air volume, for example a surge tank, which may be called an upstream end of a transmission path of the pressure wave. Then, it goes back to the intake valve, which may be called a downstream end of the pressure wave transmission path. The intake valve cyclically opens and closes synchronously with a rotation of a crankshaft of the engine. If the pressure wave reaches at the intake valve which is closed, it reflects there and goes upstream again. If a positive pressure portion of pressure wave reaches at the intake valve that is open, more air can be charged into the cylinder, which is called a supercharging effect. 
   The variable intake system can synchronize the pressure wave reaching at the intake valve with the openings of the intake valve by adjusting the effective length of the pressure wave transmission path in accordance with a frequency of the intake valve opening that corresponds to the engine speed. It can make a positive pressure portion of the pressure wave to reach at the intake valve at timing when it is open. Therefore, the variable intake system can increase the air charge through a range of the engine speed by changing the effective length. 
   On the other hand, when a desired intake airflow is smaller, in particular, at a lower engine speed, such as during an engine idling, the air charge to the individual cylinder is usually relatively small. The effective lengths of the pressure wave transmission path for the respective cylinders may vary, for example due to constraints of the intake system design. The variation of the effective lengths may cause the air charge to fluctuate between the respective cylinders, because some of the cylinders may have the supercharging effect and some may not. When the desired airflow is smaller, the air charge fluctuation may be relatively greater. It may cause variation of air-fuel ratio in the combustion chamber, for example. In turn, it may cause unstable combustion and output torque fluctuation, and increase noise or vibration of the engine. 
   Further, if the airflows to the respectively cylinders are independently controlled, for example by adjusting intake valve lifts, the individual cylinder charges may fluctuate from each other due to mechanical variations of the components, for example. When the desired airflow is smaller, the fluctuation is great relatively to the airflow. When it combines with the cylinder charge fluctuation caused by the variable intake system, the total cylinder charge fluctuation may increase significantly. Therefore, there is a need to improve the variable intake system in terms of the operation for the smaller airflow. 
   SUMMARY 
   Accordingly, there is provided, in one aspect of the present description, a method of controlling an intake air passage of an internal combustion engine. The intake air passage cyclically communicates to a combustion chamber of the internal combustion engine, thereby inducting fresh air into said combustion chamber. The cyclic communication of the intake air passage to the combustion chamber generates a pressure wave in the intake air passage. The method comprises reducing an effective length of a transmission path of the pressure wave in an upstream direction of the intake air passage as a desired airflow to the combustion chamber decreases. 
   In accordance with the method, the effective length of the pressure wave transmission path is reduced as desired air flow decreases. With the reduced effective length, the pressure wave bounces back and forth between ends of the transmission path more often until the next cyclic communication begins. The more bouncing attenuates the pressure wave more by the time of the next cyclic communication. Therefore, the cylinder air charge can be stabilized when the desired air flow is decreased. 
   In another aspect of the present invention, there is provided a method of controlling the intake air passage described above. The method comprises reducing an effective length of the transmission path of the pressure wave in the upstream direction of the intake air passage and retarding beginning of the cyclic communication as desired air flow to the combustion chamber decreases. 
   In accordance with the method, the effective length of the pressure wave transmission path is reduced as desired air flow decreases. Therefore, the pressure wave may be attenuated as described above. At the same time, the beginning of the cyclic communication of the intake air passage to the combustion chamber, for example the intake valve opening, is retarded when the desired air flow is smaller. At that time, the intake valve lift may be reduced, as described above. It may cause the fluctuation of the cylinder air charges. With the retarded beginning of the cyclic communication, the airflow into the combustion chamber has a faster rate when the communication is substantially established, for example, when the intake valve is substantially open. This may suppress the fluctuation of the air charge between the cylinders. However, at the same time, a greater negative pressure is generated just upstream of the intake valve in the intake air passage at the beginning of the communication, for example because the piston may position lower, and the cylinder pressure may be more negative, when the intake valve is substantially open. The greater negative pressure may generate a greater pressure wave in the intake air passage. But, as described above, the reduced effective length may attenuate the pressure wave. Therefore, the cylinder air charge can be stabilized when the desired airflow is smaller. 

   
     BRIEF DESCRIPTION OF THE DRAWINGS 
     The advantages described herein will be more fully understood by reading an example of embodiments in which the above aspects are used to advantage, referred to herein as the Detailed Description, with reference to the drawings wherein: 
       FIG. 1  is a schematic view showing an engine system according to an embodiment of the present description; 
       FIG. 2  shows an operation of a shutter valve of a variable intake system with an open state (A) and an closed state (B) of the shutter valve in accordance with the embodiment; 
       FIG. 3  shows a perspective view of an intake valve drive mechanism including a variable cam timing mechanism and a variable valve lift mechanism in accordance with the embodiment; 
       FIG. 4  shows a side view of the variable valve lift mechanism for a valve open state ( 1 ) and a valve closed state (B) with a greater valve lift in accordance with the embodiment; 
       FIG. 5  shows a side view of the variable valve lift mechanism for a valve open state ( 1 ) and a valve closed state (B) with a smaller valve lift in accordance with the embodiment; 
       FIG. 6  is explanatory diagrams for the greater valve lift (A) and the smaller valve lift (B) respectively illustrated in  FIGS. 4 and 5 ; 
       FIG. 7  shows various valve lift profiles generated by the valve lift mechanism in accordance with the embodiment; 
       FIG. 8  is a flowchart showing one of the control routines executed by an engine controller in accordance with the embodiment; 
       FIG. 9  is a flowchart showing one of the control routines executed by the engine controller in accordance with the embodiment; 
       FIG. 10  is a flowchart showing one of the control routines executed by the engine controller in accordance with the embodiment; 
       FIG. 11  shows a profile of a control angle of the variable valve lift mechanism over a range of engine operating condition in accordance with the embodiment; 
       FIG. 12  shows a profile of a control angle of the variable cam timing mechanism over the range of engine operating condition in accordance with the embodiment; 
       FIG. 13  shows various valve lift profiles generated by a combination of the variable cam timing mechanism and the variable valve lift mechanism in accordance with the embodiment; 
       FIG. 14  shows graphs of pressures in an intake ports in accordance with the embodiment (shown by a solid line in (A)) and comparative examples (shown by a dotted line in (A) and solid and dotted lines in (B)); and 
       FIG. 15  shows open and closed states of the shutter valve with a line of most advance cam timing over the range of engine operating condition in accordance with the embodiment. 
   

   DETAILED DESCRIPTION 
   An embodiment of the present description will now be described with reference to the drawings, starting with  FIG. 1 , which shows a schematic view of an engine system mounted on a vehicle, such as an automotive vehicle. The output of the engine system is transmitted to vehicle driving wheels through a power transmission mechanism as is well known in the art. The engine system comprises an internal combustion engine  1 , an engine controller  100  that controls the engine  1  and other subsystems described below. 
   The internal combustion engine  1  has four cylinders  2 , in the present embodiment, although only one cylinder is shown in  FIG. 1  and the engine  1  may have any number of cylinders. The engine  1  comprises a cylinder block  3 , and a cylinder head  4 , which is arranged on the cylinder block  3 . The cylinder  2  accommodates a piston  5  which slides therein. As is well known in the art, the cylinder block  3  rotationally supports a crankshaft  6  using journals, bearings and the like. Further, a connecting rod  7  links the crankshaft  6  and the piston  5 . The cylinder head  4 , the cylinder  2 , and the piston  5  collectively define a combustion chamber  8  inside. 
   Although only one is illustrated in  FIG. 1 , two intake ports  9  are formed in the cylinder head  4 , and respectively open to the combustion chamber  8 . Likewise, two exhaust ports  10  are formed in the cylinder head  4 , and respectively open to the combustion chamber  8 . Intake valves  11  and exhaust valves  12  are respectively capable of shutting the intake ports  9  and the exhaust ports  10  from the combustion chamber  8  as shown in  FIG. 1 . Valve drive mechanism  101  and  102  respectively drives the intake and exhaust valves  11  and  12  as described in greater detail below. 
   A spark plug  15  for the each cylinder  2  is mounted to the cylinder head  4  in the well known manner such as threading. An ignition circuit or system  16  receives a control signal SA from an engine controller  100 , and provides electric current to the spark plug  15  so that it makes a spark at the plug gap in the combustion chamber  8  at desired ignition timing. 
   For the each cylinder  2 , a fuel injector  17  is mounted to the cylinder head  4  at one side of a cylinder center axis in a known manner such as using a mounting bracket. A tip end of the injector  17  faces the inside of the combustion chamber  8  from a space vertically below and horizontally between the two intake ports  9 . A fuel supply system  18  includes a high pressure pump and an injector driver circuit not shown, and supplies fuel, in this case gasoline, from a fuel tank not shown. Also, the fuel supply system  18 , particularly an injector driver circuit therein, activates a solenoid of the injector  17  to open the spray nozzles in accordance with a control signal FP from the engine controller  100 , in order to inject desired amount of fuel at desired timing. The fuel is not limited to the gasoline, but may be any fuel including ethanol and hydrogen as far as it can be ignited by the spark from the spark plug  15 . The injector  17  is not limited to being arranged to directly inject fuel into the combustion chamber  8  (direct fuel injection), but it may be arranged to inject fuel into the intake port  9  (port fuel injection). 
   An intake manifold  30  is attached to the cylinder head  4  in well known manner such as by threading with bolts. Each of the intake ports  9  connect in fluid communication to a surge tank  31  (only shown in  FIG. 2 ) through upstream and downstream branch intake passages  32  and  33  all of which are integrally formed with the intake manifold  30 . The branch intake passages  32  and  33  are provided for the respective cylinders  2  or combustion chamber  8 . 
   Therefore in this embodiment having four cylinders, there are four sets of the branch intake passages  32  and  33 . They are connected in fluid communication to each other by a communication chamber  34  at respective connecting portions between the upstream and downstream branch intake passages  32  and  33 . 
   Shutter valves  35  are arranged between the connecting portions of the branch intake passages  32  and the communication chamber  34 , which are pivoted around a common axis and capable of opening and closing the fluid communication between the branch intake passages  32  and  33  and the communication chamber. A shutter actuator  36  is provided to operate the shutter valves  35  to open and close the fluid communication in accordance with a signal SV from the engine controller  100 . 
   Upstream of the surge tank  31 , a throttle body  20  is coupled to the intake manifold  30 . It accommodates a throttle valve  21  therein. It pivots and regulates engine intake airflow to the surge tank  31  from an air cleaner not shown, as is well known in the art. A throttle actuator  22 , for example, an electric motor or a vacuum actuator, adjusts an opening of the throttle valve  21  in accordance with a control signal TVO from the engine controller  100 . 
   The exhaust ports  10  are connected to an exhaust manifold  23 . Exhaust gas from the combustion chamber  8  flows through the exhaust manifold  23  to a catalytic converter  24 , and then to an exhaust pipe not shown, as is well known in the art. 
   Referring to  FIG. 2 , the supercharging effect by an intake air passage consisting of the intake port  9  and the branch intake passages  32  and  33  will now be described. There are shown two different states of the intake air passage, namely a closed state (A) and an open state (B) of the shutter valves  35 . 
   In  FIG. 2(A) , the shutter valves  35  shut off the communication chamber  34  from the branch intake passages  32  and  33 . When the intake valve  11  opens the communication of the intake air passage to the combustion chamber  8  at the beginning of an intake stroke of a cylinder cycle, the descending movement of the piston  5  causes a negative pressure in a throat  9   a  of the intake port  9  just upstream of the intake valve  11 . The negative pressure wave transmits through the intake air passage, i.e. the intake port  9  and the branch intake passages  32  and  33 , to the surge tank  31 . Since volume of the surge tank  31  is much greater than volume of the branch intake passages  32  and  33 , the pressure wave is reflected at the upstream end of the upstream branch intake passages  33  at the surge tank, and at the same time, it turns its phase in reverse, in other words, to be the positive pressure wave. Then, it goes back through the branch intake passages  32  and  33  to the throat  9   a  of the intake port  9 . 
   If the positive pressure wave reaches the intake port throat  9   a  when the intake valve  11  is open, or in other words, a synchronization between the pressure wave and the intake valve timing occurs, it boosts cylinder air charge, which is so-called inertia supercharging effect. The synchronization occurs if the sound speed times cycle time of the intake valve operation equals to an effective length for which the pressure wave transmits back and forth from the intake valve  11 . The effective length can be calculated from an actual length of and respective sectional areas on a path of the transmission of the pressure wave. In this embodiment, the intake port  9  and the branch intake passages  32  and  33  are configured so that the synchronization occurs at a certain engine speed, for example, between 3000 and 4000 rpm. 
   If the positive pressure wave reaches the intake port throat  9   a  while the intake valve  11  is closed, the positive pressure wave reflects and turns its phase to be a negative pressure wave at the intake port throat  9   a . The negative pressure wave bounces back from the surge tank  31  to keep the pressure oscillation. As a result, this pressure oscillation remains in the branch intake passage  5  or, in other words, makes pulsation while it attenuates over time. The remaining positive pressure may still boost the intake air charge if it reaches at the intake port throat  9   a  when the intake valve  11  opens. This may occur at engine speeds less than the certain synchronizing engine speed and may be called a pulsation effect. 
   On the other hand, when the shutter valve  35  opens the branch intake passages  32  to the communication chamber  34  as shown in  FIG. 2(B) , the negative pressure wave from the intake port throat  9   a  reaches the communication chamber  34  prior to reaching the surge tank  31  because the communication chamber  34  is closer to the intake port throat  9   a . Since volume of the communication chamber  34  is much greater than that of the branch intake passage, the negative pressure reflects at the communication chamber  34  rather than at the surge tank  31 . 
   Therefore, time for the pressure wave transmitting between the intake port throat  9   a  and the reflecting point is shorter in the case of the shutter valve  35  open than closed. In other words, the pressure wave transmits for a shorter distance or the reduced effective length of the pressure wave transmission. Consequently, the inertial supercharging effect occurs at higher engine speeds than the synchronizing engine speed for the closed shutter valve  35 . 
   Although a wall  37  divides the surge tank  31  and the communication chamber  34  in the above embodiment, alternatively, the wall  37  may be removed and the communication chamber  34  may be part of the surge tank  31 . 
   Referring to  FIG. 3 , the valve drive mechanism  101  for the intake valves  11  will now be described in more detail. Referring to  FIG. 3 , there is shown the valve drive mechanism  101  for the intake valves  41 . The valve drive mechanism  102  for the exhaust valves  12  has a same construction as for the intake in the present embodiment. Therefore the specific description for the mechanism  102  will be omitted. Alternatively, the valve drive mechanism  102  for the exhaust valves may be of a conventional overhead camshaft (OHC) type. The OHC type valve drive mechanism comprises a cam for pushing a valve stem, a camshaft integrally forming the cam, and a camshaft drive-train such as chain and sprocket for transmitting rotational movement of the crankshaft  6  to the camshaft, as is well known in the art. 
   The valve drive mechanism  101  has a variable cam timing (VCT) mechanism  103 , which is linked to the crankshaft  21  through a chain drive mechanism including a driven sprocket  104 , a drive sprocket at the crankshaft  21 , and a chain not shown and engagingly wounded around the drive and driven sprockets. The VCT mechanism  103  comprises a casing, which is affixed to the sprocket  103  to rotate with it, and a rotor, which is affixed to an inner shaft  105  and rotates with it. Between the casing and the rotor of the VCT mechanism  103 , there are formed a plurality of hydraulic chambers, which are circumferentially arranged around the rotational axis X. Fluid pressurized by a pump, such as engine oil, is selectively supplied to each of the hydraulic chambers to make a pressure difference between the opposing chambers. A VCT control system  201  including an electromagnetic valve  106  adjusts the hydraulic fluid supplied to the chambers. The electromagnetic valve  106  cyclically switches hydraulic acting directions to the chambers in a duty ratio in accordance with a control signal θ VCT  from the engine controller  100  and an actual phase difference between the sprocket  104  and the inner shaft  105 , thereby achieving a desired rotational phase of the inner shaft  105 , as is known in the art. 
   The inner shaft  105  has an eccentric disc-shaped cam  106  for each of the cylinders  22 . The eccentric cam  106  is formed integrally but not coaxially with the inner shaft  105  and rotates at a phase defined by the VCT mechanism  103 . Freely rotationally fitted around the eccentric disc  106  is an inner surface of a ring arm  107 . Therefore, the ring arm  107  can self rotate around a center axis Y of the eccentric cam  106  (only shown in  FIG. 6 ) and orbit around the rotational axis X, as the inner shaft  105  rotates around the rotational axis X. 
   Arranged around the inner shaft  105  is a rocker connector  110  for each of the cylinders  22 . The rocker connector  110  pivots coaxially with the inner shaft  105 , in other words, around the axis X, and integrally forms first and second rocker cams  111  and  112 . The rocker connector  110  forms a bearing journal at its outer circumferential surface, so that a bearing cap not shown arranged on the cylinder head  24  can rotationally support the rocker cam parts  110  through  112 . As shown in  FIG. 4 , each of the rocker cams  111  and  112  has a cam surface  111   a  and a basic circular surface  111   b , either of which contacts to an upper surface of a tappet  115 , as a conventional valve drive cam does, except that the rocker cams do not continuously rotate, but rocks. The tappet  115  is supported by a valve spring  116 , which is sustained between retainers  117  and  118 , as is known in the art. 
   Referring back to  FIG. 3 , arranged above and in parallel with the assembly of inner shaft  105  and the rocker cam parts  110  through  112  is a control shaft  120 , which is rotationally supported by bearings not shown. The control shaft  120  integrally forms a worm gear  121  coaxially at its outer peripherally. The worm gear  121  engages with a worm  122 , which is affixed to an output shaft of an electric motor  123  that is controlled by an VVL control system  202  shown in  FIG. 1 . Therefore, the motor  123  may rotate the control shaft  120  to its desired position, in accordance with a control signal θ VVL  from the engine controller  15 . 
   Four control arms  131  for the respective cylinders  22  are attached to the control shaft  120 , so that the control arms  131  can pivot integrally with the control shaft  120 . A control link  132  couples each of the control arms  131  and the respective ring arm  107  through a control pivot  133  and a common pivot  134 . Then, a rocker link  135  couples the ring arm  107  and the first cam  111  through the common pivot  134  and a rocking pivot  136 . 
     FIG. 4  and  FIG. 6(A)  show a condition where a valve lift is greater. The control arm  131  is adjusted to define a VVL control angle θ VVL     —     A  between the horizontal plane shown by a dotted line in  FIG. 6(A)  and a line connecting the center axes of the control shaft  120  and the control pivot  133 . 
   When the inner shaft  105  rotates around the axis X clockwise on the sheet of Figures from a no-lift state ( 1 ) to a maximum-lift state ( 2 ) in  FIG. 4  or from a state shown by broken lines to a state shown by solid lines in  FIG. 6(A) , the common center Y of the eccentric cam  106  and the ring arm  107  orbits clockwise from points Y 1A  to Y 2A  around the axis X as shown in  FIG. 6(A) . The orbital movement of the ring arm  107  causes a rocking movement of the control link  132  by an angle θ 132A  around the control pivot  132  due to a first four-link relationship consisting of four pivots X, Y,  133  and  134  and the corresponding links. Therefore, the common pivot  134  rocks around the control pivot  133 . The common pivot  134  is at its rotational end positions when the axis X, the common center Y and the common pivot  134  are in line. One of the end positions of the common pivot  134  is shown by the solid lines in  FIG. 6(A) . 
   Four pivots  133 ,  134 ,  136  and X and corresponding links consist a second four-link relationship. It converts the rocking movement of the common pivot  134  by the angle θ 132A  to a rocking movement of the rocker cam  111  or  112  by an angle θ 111A  around the axis X. When the common center Y is located at Y A1 , the cam  111  is at one of its angular end positions because the common pivot is at its rotational end as described above and as shown in  FIG. 6(A) . 
   When the cam surface  111   a  of the rocker cam  111  or  112  contacts the tappet top surface  115   a  as in the state ( 1 ) of  FIG. 4  and as shown by the solid line in  FIG. 6(A) , the rocker cam  111  or  112  moves down the tappet  115  against the valve spring  116 . Then, the tappet  115  causes the intake valve  41  to move down to its maximum valve lift under the angle θ VVL     —     A  of the control arm  131  in  FIG. 6(A) . 
   On the other hand, when the basic circular surface  111   b  contacts the tappet top surface  115   a  as shown in the state ( 1 ) of  FIG. 4  and by the broken line in  FIG. 6(A) , the tappet  115  is not pushed down, because the basic circular surface  111   b  has a constant radius smaller than a distance between a point of the cam surface  111   a  and the axis X. Therefore, the angle θ VVL     —     A  or the angular position of the control arm  131  causes a valve lift h A  as shown in  FIG. 6(A) . 
     FIGS. 5 and 6(B)  show a condition of smaller valve lift L B . The control arm  131  is adjusted to define an angle θ 131B  between the horizontal plane shown by the dotted line and the line connecting the center axes of the control shaft  120  and the control pivot  133  as shown in  FIG. 6(B) . In this Figure, as the inner shaft  105  rotates clockwise, the common center Y orbits from points Y 1B  to Y 2B . For the illustrative purpose, the point Y 1B  is the same point as Y 1A  in  FIG. 6(A) . The position Y 2B  is one of angular end positions where the axis X, the common center Y and the common pivot  133  are in line. 
   The first four-link relationship consisting of the pivots X, Y,  133  and  134  and the others causes an angular movement of the control link  132  by an angle θ 132B . Then, the second four-link relationship consisting of the pivots  133 ,  134 ,  136  and X converts the angular movement of the control link  132  or the common pivot  134  into a rocking movement of the rocking cam  111  or  112  with an angle θ 111B . When the common center Y is located at Y B1 , the cam  111  is at one of its angular end positions because the common pivot Y is at its rotational end as described above and as shown in  FIG. 6(B) . 
   When the basic circular surface  111   b  contacts tappet top surface  115   a  as shown in the state ( 1 ) of  FIG. 5  and by the broken line in  FIG. 6(B) , the tappet  115  is not pushed down as in the case of  FIG. 6(A) . When the cam  111  is positioned as shown by the solid line in  FIG. 6(B) , the cam surface  111   a  contacts the tappet top surface  115   a  and pushes down the tappet  115  most under the angular position θ 131B  of the control arm  131 . As can be seen from  FIG. 6 , a valve lift h B  is much smaller than the valve lift h A . Therefore, as the angle θ VVL  is smaller, the peak valve lift h decreases. If the angle θ VVL  is further increased, the valve lift can be zero depending on the configuration of a variable valve lift (VVL) mechanism. 
   Further, as the angle θ VVL  is smaller, the rocking angle θ 111  decreases, and the angular position Y 2  of the common center Y, with which the maximum valve lift is obtained, shifts counterclockwise. These can be seen from valve lift curves in  FIG. 7 . A valve lift curve L A  illustrates the greater valve lift state with the angle θ VVL     —     A  shown in  FIGS. 4 and 6(A) , and a valve lift curve L B  illustrates the smaller valve lift state with the angle θ VVL     —     B  shown in  FIGS. 5 and 6(B) , for a case where only the VVL actuator  123  is operated with the VCT mechanism  103  setting the inner shaft  105  at a fixed angular phase with respect to the crankshaft  6 . 
   As can be seen from  FIG. 7 , the variable valve lift (VVL) mechanism has characteristics where valve opening duration increases, peak valve lift timing is retarded and valve closing timing is retarded as the maximum valve lift increases. Further it can be seen that the valve opening timing does not change so much as the valve closing timing does. 
   This valve lift characteristic is preferable for regulating air charge inducted into the combustion chamber  8 . When the throttle valve  21  is closed to regulate the air charge, it causes restriction of intake air flow to the combustion chamber  8 , and the kinetic energy of the engine moving parts, such as the piston  5  and the crankshaft  6 , are spent for pumping in the restricted air in an intake stroke of an engine cylinder cycle. This is called “pumping loss”. 
   Rather, the valve lift characteristic shown in  FIG. 7  can regulate air charge with less throttling and less pumping loss. 
   Basically, the air charge will be decreased as the intake valve closing timing is advanced or retarded from certain timing. The certain timing is at the bottom dead center of the piston if the engine speed is extremely low because there is no inertia of the intake airflow. Practically, it retards as the inertia of the intake airflow increases. The inertia more heavily weights on the intake airflow rate or engine speed. Further, greater valve lift is required for greater airflow. 
   Otherwise, flow restriction may occur at the intake port throat  9   a  and the intake valve  11  when the air flow increases in dependence on the increased airflow rate or air charge. The VVL mechanism described above has the characteristic where the valve closing timing is retarded as the valve lift is greater as shown in  FIG. 7  and described above. Therefore, it can preferably meet to the requirement for regulating air charge into the combustion chamber  8  with less throttling. 
   The engine controller  100  is a microcomputer based controller having a central processing unit which runs programs using data, memories, such as RAM and ROM, storing the programs and data, and input/output (I/O) bus inputting and outputting electric signals, as is well known in the art. The engine controller  100  receives signals from various sensors. As shown in  FIG. 1 , the input signals to the engine controller  100  include a signal AF from an air flow meter  51  arranged in the air cleaner described above and known in the art, a pulse signal from a crank angle sensor  52  based on which an engine speed N E  is computed by the engine controller  100 , a signal a from an accelerator position sensor  53  detecting a position of an accelerator pedal  54 , a signal MAP from a pressure sensor  55  detecting a pressure in the intake manifold  30 , and an a signal EGO from an oxygen sensor  56  detecting an oxygen concentration in the exhaust gas upstream of the catalytic converter  24 . 
   Based on these input signals, the engine controller  100  computes and outputs various control signals including the signal SA to the ignition system  16 , the signal FP to the fuel system  18 , the signal TVO to the throttle actuator  22 , the signal SV to the shutter actuator  36 , a signal θ VCT  to the VCT control system  201 , and a signal θ VVL  to the VVL control system  202 . 
   Control routines R 1  through R 3  executed by the controller  100  will now be described with reference to flow charts of  FIGS. 8 through 10 . The illustrated routines R 1  through R 3  are mainly for the shutter actuator  36 , the VCT control system  201  and the VVL control system  202 . The engine controller  100  executes the rest of control on the engine  1  in manners known in the art. 
   For example, the fuel signal FP may be computed based on the intake airflow AF detected by the airflow meter  31  and the engine speed N E  so as to achieve a target air-fuel ratio in the combustion chamber and further corrected based on the oxygen concentration in the exhaust gas detected by the exhaust gas oxygen sensor  56 . The intake airflow AF is controlled by using the throttle valve  20  or the intake valve drive mechanism  101  based on desired engine output torque TQ D  and the engine speed N E . Therefore, if the air-fuel ratio is constant, most likely to be a stoichiometric air fuel ratio for a fuel supplied to the engine  1 , the intake airflow AF is a function of a product of the desired engine output torque TQ D  and the engine speed N E . 
   As shown in  FIG. 8 , the routine R 1  determines desired engine output torque TQ D  based on the position α of the accelerator pedal  34  and the engine speed N E . After the start, the routine R 1  proceeds to a step S 11  and reads signals, which the engine controller  100  has read from the sensors described above and stored in its memory. The read signals include the position a of the accelerator pedal  34  detected by the accelerator position sensor  33  and the engine speed N E  that the engine controller  100  computes from the crank angle signal detected by the crank angle sensor  52 . 
   Then, the routine R 1  proceeds to a step S 12  and determines a desired engine output torque TQ D  based at least on the accelerator position α and the engine speed N E  by referring to a TQ D  map. The desired torque TQ D  is mapped to increase in proportion to increase of the accelerator position α and the engine speed N E . Additionally, other engine parameters such as a transmission gear ratio, a cruise control signal, engine temperature, ambient temperature, or ambient pressure may be considered for the desired torque determination. After the step S 12 , the routine R 1  proceeds to a step S 13 , and the engine controller  100  stores the desired engine output torque TQ D  determined at the step S 12  into its memory. Then, it returns. 
   As shown in  FIG. 9 , the routine R 2  controls the intake valve drive mechanism  101 . After the start, the routine R 2  proceeds to a step S 21  and reads the engine speed N E  and the desired engine output torque TQ D  that is determined at the step S 12  and stored in the engine controller memory at the step S 13 . Then, the routine R 2  proceeds to a step S 22  and determines a VVL control angle θ VVL  based on the engine speed N E  and the desired engine output torque TQ D  by referring to a θ VVL  map. The VVL control angle θ VVL  indicates an angular position of the control shaft  120  of the VVL mechanism shown in  FIGS. 3 through 6 . It corresponds to a peak valve lift as described above. Although it will be described in more detail later with reference to  FIG. 11 , the VVL control angle θ VVL  is mapped so that peak valve lifts are greater as the engine speed N E  increases or the desired torque TQ D  increases. After the VVL control angle θ VVL  is determined, the routine R 2  proceeds to a step S 23 , and the engine controller  100  outputs the signal θ VVL  to the VVL control system  202  to control the VVL actuator  123  to adjust the control arm  120  at the position corresponding to the signal θ VVL . 
   Then, the routine R 2  proceeds to a step S 24  and determines a VCT control angle θ VCT  based on the engine speed N E  and the desired engine output torque TQ D  by referring to a θ VCT  map. The VCT control angle θ VCT  indicates a relative phase between the crankshaft  6  and the inner shaft  105  of the intake valve drive mechanism  101  shown in  FIG. 3 . Although it will be described in more detail later with reference to  FIG. 12 , the VCT control angle θ VCT  is mapped in first and second areas. In the first area where the engine speed N E  and the desired torque TQ D  are relatively small, the VCT control angle θ VCT  increases so that the angular phase of the inner shaft  105  gets more advanced as the engine speed N E  increases or the desired torque TQ D  increases. In the second area where the engine speed N E  and the desired torque TQ D  are relatively great, the VCT control angle θ VCT  decreases so that the angular phase of the inner shaft  105  gets more retarded as the engine speed N E  increases or the desired torque TQ D  increases. After the VCT control angle θ VCT  is determined, the routine R 2  proceeds to a step S 25 , and the engine controller  100  outputs the signal θ VCT  to the VCT control system  201  to control the VCT actuator  103  to adjust the inner shaft  105  to the angular phase corresponding to θ VCT . Then, the routine R 2  returns. 
   As shown in  FIG. 10 , the routine R 3  controls the shutter valve  35  based on the desired torque TQ D  and the engine speed N E . After a start, the routine R 3  proceeds to a step S 31  and reads the engine speed N E , the desired engine output torque TQ D  that is determined at the step S 12  and stored in the engine controller memory at the step S 13 , and the signal SV for the shutter actuator  36  that is determined and stored in the engine controller memory during the execution of this routine R 3 . Then, it proceeds to a step S 32  and determines whether or not the read engine speed N E  is greater than a first upper predetermined speed N E1U  which may be for example 2500 rpm. When it is determined that the engine speed N E  is greater than the first upper predetermined speed N E1  (YES) at the step S 32 , the routine proceeds to a step S 33 , and the engine controller  100  outputs a signal SV OPEN  to the shutter valve actuator  36  to open the shutter valve  35 . Then, the routine R 3  returns. 
   On the other hand, when it is determined that the engine speed N E  is not greater than the first upper predetermined speed N E1U  (NO) at the step S 32 , the routine proceeds to a step S 34  and determines whether or not the read engine speed N E  is greater than a first lower predetermined speed NELL which is set slightly lower than N E1U  for avoiding excessively frequent actuations of the shutter actuator  36 . When it is determined that the engine speed N E  is greater than the first lower predetermined speed NELL (YES) at the step S 34 , the routine proceeds to a step S 35  and determines whether or not the shutter valve  35  is currently open based on the signal SV read at the step S 31 . When it is determined that the shutter valve  35  is open (YES) at the step S 35 , the routine proceeds to the step S 33 , and the engine controller  100  outputs a signal SV OPEN  to the shutter valve actuator  36  to keep the shutter valve  35  open. On the other hand, it is determined that the shutter valve  35  is closed (NO) at the step S 35 , the routine proceeds to a step S 36 , and the engine controller  100  outputs a signal SV OPEN  to the shutter valve actuator  36  to keep the shutter valve  35  closed. Therefore, when the engine speed N E  is between the first upper and lower predetermined speeds N E1U  and N E1L , the shutter valve  35  will keep its current open or closed state. After either of the steps S 33  and S 36 , the routine returns. 
   When it is determined that the engine speed N E  is not greater than the first lower predetermined speed N E1L  (NO) at the step S 34 , the routine proceeds to a step  37  and determines whether or not the read engine speed N E  is greater than a second upper predetermined speed N E2U  which may be for example 1000 rpm. When it is determined the engine speed N E  is greater than the second upper predetermined speed N E2U  (YES) at the step S 37 , the routine R 3  proceeds to the step S 36 , and the engine controller  100  outputs the signal SV CLOSE  to the shutter actuator  36  to close the shutter valve  35 . Then, the routine R 3  returns. 
   When it is determined the engine speed N E  is not greater than the second upper predetermined speed N E2U  (NO) at the step S 37 , the routine R 3  proceeds to a step S 38  and determines whether or not the read engine speed N E  is greater than a second lower predetermined speed N E2L  which is set slightly lower than N E2U . When it is determined that the engine speed N E  is greater than the second lower predetermined speed N E2L  (YES) at the step S 38 , the routine R 3  proceeds to the step S 35  and then the step S 33  or S 36 , and the engine controller  100  sends the signal SV OPEN  or SV CLOSE  to keep the shutter valve  35  at its current open or closed state as described above. Then, the routine R 3  returns. 
   When it is determined that the engine speed N E  is not greater than the second lower predetermined speed N E2L  (NO) at the step S 38 , the routine R 3  proceeds to a step S 39  and determines whether or not the read desired engine torque TQ D  is greater than a first upper predetermined engine torque TQ 1U . When it is determined that the desired engine torque TQ D  is greater than the first upper predetermined engine torque TQ 1  (YES) at the step S 39 , the routine proceeds to the step S 36 , and the engine controller  100  outputs the control signal SV CLOSE  to the shutter actuator  36  to close the shutter valve  35 . Otherwise, the routine proceeds to the step S 40 , and determines whether or not the read desired engine torque TQ D  is greater than a first lower predetermined engine torque TQ 1L . 
   When it is determined that the desired engine torque TQ D  is not greater than the first lower predetermined engine torque TQ 1L  (NO) at the step S 40 , the routine R 3  proceeds to the step S 33 , and the engine controller  100  outputs the control signal SV OPEN  to the shutter actuator  36  to open the shutter valve  35 . 
   On the other hand, when it is determined that the desired engine torque TQ D  is greater than the first lower predetermined engine torque TQ 1L  (YES) at the step S 40 , the routine R 3  proceeds to the step S 35  and then the step S 33  or S 36 , and the engine controller  100  sends the signal SV OPEN  or SV CLOSE  to keep the shutter valve  35  at its current open or closed state as described above. Then, the routine R 3  returns. 
     FIG. 11  illustrates in more detail the θ VVL  map which is referenced at the step S 22  of the routine R 2  shown in  FIG. 9  by indicating peak valve lift that corresponds to the VVL control angle θ VVL . It consistently increases as the engine speed N E  increases or the desired engine torque TQ D  increases. As described above, since the intake airflow is the function of the product of the desired engine torque TQ D  and the engine speed N E , the VVL control angle θ VVL  consistently increases as the desired intake airflow to the combustion chamber  8  increases. 
   Because of the valve lift characteristics of the VVL mechanism shown in  FIG. 7  and described above, if the VCT control angle θ VCT , which is usually adjusted at the step S 24 , is fixed, the peak valve lift increases and the intake valve closing timing is retarded, as shown in  FIG. 11 . 
     FIG. 12  illustrates in more detail the θ VCT  map which is referenced at the step S 24  of the routine R 2  shown in  FIG. 9  by indicating a relative phase angle of the inner shaft  105  with respect to the phase of the driven sprocket  104  shown in  FIG. 4 . In the first engine operating area of higher torque TQ D  and higher engine speed N E , which is the upper right side of a line B in  FIG. 12 , the VCT control angle θ VCT  is consistently retarded as the engine speed N E  increases, the desired engine torque TQ D  increases, or the desired intake airflow to the combustion chamber  8  increases. On the other hand, in the second engine operating area of lower torque TQ D  and lower engine speed N E , which is the lower left side of the line B in  FIG. 12 , the VCT control angle θ VCT  is consistently retarded as the engine speed N E  decreases, the desired engine torque TQ D  decreases, or the desired intake airflow to the combustion chamber  8  increases. 
   The control routines described above, specifically the routine R 2 , uses both of the θ VVL  map and the θ VCT  map. In the first engine operating area of  FIG. 12 , as the engine speed N E , the desired engine torque TQ D , or the desired intake airflow to the combustion chamber  8  increases, the both VVL control angle θ VVL  and VCT control angle θ VCT  increase. This more greatly retards the intake valve closing timing, since the each angle increase retards the intake valve closing timing. Therefore, the air charge to the combustion chamber  8  can be more effectively regulated with less throttling. 
   On the other hand, in the second engine operating area of  FIG. 12 , as the engine speed N E , the desired engine torque TQ D , or the desired intake airflow to the combustion chamber  8  increases, the VVL control angle θ VVL  increases but the VCT control angle θ VCT  decreases. Then, the closing timing advance caused by VCT control angle decrease compensates the closing timing retard caused by the VVL control angle increase. Therefore, the intake valve closing timing does not change so much as in the first engine operating area. 
   Instead, the intake valve opening timing advances as the engine speed N E  decreases, the desired engine torque TQ D , the desired intake airflow decreases, because the change of the VVL control angle θ VVL  less affects the intake valve opening timing as shown in  FIG. 7 . 
     FIG. 13  shows various lift curves L 1  through L 4  of the intake valve  11  that are generated by the control routine R 2 . The lift curve L 1  is created at a lower engine torque and speed condition in the second engine operating area in  FIG. 12 . The lift curve L 2  is generated on the boundary between the first and second engine operating conditions. The lift curve L 3  is generated at a higher engine torque and lower engine speed condition in the first engine operating condition and corresponds to the peak valve lift equal to 5 mm shown in  FIG. 11 . 
   Finally, the lift curve L 4  is created at a higher engine torque and speed condition in the first engine operating condition. 
   Between the lift curves L 2  and L 4 , or in other words, within the first engine operating area shown in  FIG. 12 , the valve closing timing is retarded as the peak valve lift increases. Between the lift curves L 2  and L 3 , the intake valve  11  is substantially closed before BDC. As described above, intake valve closing timing for the greatest air charge varies in dependence on the engine speed. The closing timing of the lift curve L 3  corresponds to the greatest air charge in the mid engine speed range. The closing timing of the lift curve L 2  corresponds to the smaller air charge in the mid engine speed range. Therefore, the air charge can be regulated in the mid engine speed range by setting the valve lift curve between the curves L 2  and L 3 . 
   Between the lift curves L 3  and L 4 , the intake valve  11  is substantially closed before BDC. In the higher engine speed range, the closing timing of the lift curve L 3  corresponds to the smaller air charge due to the inertia of the intake airflow described above, and that of L 4  corresponds to the greatest air charge. Therefore, the air charge can be regulated in the higher engine speed range by setting the valve lift curve between the curves L 2  and L 3 . 
   On the other hand, between the lift curves L 1  and L 2 , in other words, within the second engine operating condition shown in  FIG. 12 , the intake valve opening timing is retarded while the closing timing does not substantially change as the peak valve lift decreases. On the lift curve L 1 , the intake valve  11  starts opening after the top dead center (TDC) when the piston  8  starts descending as shown in  FIG. 13 . Then, the lift curve L 1  reaches at the peak lift in the latter half of the piston descending stroke. Even with the smaller peak valve lift, the greater piston descending movement in the latter half of the piston descending stroke causes greater airflow. Even with some variation of the intake airflow between the cylinders, such as geometric variation of the intake passages, the greater airflow makes the amount of flow to be more consistent between the cylinders. Therefore, especially in a lower speed and torque condition where the desired air charge is relatively small and its deviation may affect more the engine operating condition such as an air-fuel ratio in the combustion chamber, the retarded intake valve opening timing of the valve lift L 1  makes the engine operation more stable. 
   With the retarded intake valve opening timing of the lift curve L 1 , the piston  5  has already descent for some distance and generates negative pressure within the combustion chamber  8  when the intake valve  11  is substantially opened.  FIG. 14  shows a comparison between pressure changes with the intake valve opening timings after TDC (for the lift curve L 1 ) (A) and before TDC (B) at the intake port throat  9   a  under a same engine operating condition where the engine speed and air charge are relatively low such as in an engine idle condition. As can be seen from  FIG. 14 , the retarded intake valve opening timing of the lift curve L 1  generates greater amplitude of the pressure wave. 
   Referring back to  FIG. 13 , as the lift curve moves from L 1  to L 2 , the intake valve opening timing is advanced while the closing timing does not substantially change and regulate the air charge into the combustion chamber  8 . Instead, the change of the peak valve lift may regulate the air charge. Additionally, the throttle valve  21  may be used for the air charge regulation. 
   The control routine R 3  shown in  FIG. 10  opens and closes the shutter valve  35  depending on the engine speed N E  and the desired torque TQ D  as shown in  FIG. 15 . Note that in  FIG. 15  and the following description, the first upper and lower predetermined speed N E1U  and N E1L , the second upper and lower predetermined speed N E2U  and N E2L  and the first upper and second predetermined desired torque TQ D1U  and TQ D2L  are collectively called the first predetermined speed N E1 , the second predetermined speed N E2 , and the first predetermined desired torque TQ D1 , respectively. At a higher engine speed, specifically when it is determined the engine speed N E  is greater than the first predetermined speed N E1  at the step S 32 , the shutter valve  35  is opened, and the distance between the intake valve  11  and a larger volume, in this case the communication chamber  34 , is shorter. As described above, the opening of the intake valve  11  generates the pressure wave which transmits in the intake air passage. When the shutter valve  11  is open, the effective length of the pressure wave transmission path is shorter than in the case of the closed shutter valve. In the higher engine speed range, the shorter path may cause the pressure wave to return the intake port throat  9   a  during the intake valve opening period and increase the air charge into the combustion chamber  8 . 
   On the other hand, when the shutter valve  55  is closed, the effective length of the pressure wave transmission path between the intake valve  11  and a larger volume, in this case the surge tank  31 , is longer. In the lower engine speed range, time duration for the pressure wave to take to return to the intake port throat  9   a  is longer than in the higher engine speed range. Therefore, the air charge into the combustion chamber  8  can be increased in the both higher and lower engine speed ranges by opening and closing the shutter valve  35  respectively above and below the engine speed N E1 . 
   When the engine  1  operates in an engine idle condition, or when the engine speed is not greater than the second predetermined speed N E2  as determined at the step S 38  and the desired engine torque TQ D  is not greater than the first predetermined torque TQ D1  as determined at the step S 40  of the routine R 3  shown in  FIG. 9 , the desired intake airflow is relatively small because the desired intake airflow is the function of the product of the engine speed N E  and the desired torque TQ D  as described above. In this operating area, the amplitude of the pressure wave generated by closing the shutter valve  35  may be great relatively to the amount of airflow as shown by dotted lines in  FIG. 14 . That relatively large pressure wave may affect each of the cylinders differently from each other, for example, because of the geometric variation of the intake passages. It may cause relatively great deviation between the cylinder air charges or cylinder air fuel ratios. These deviations may reduce stability of combustion of air-fuel mixture and cause fluctuation of engine speed and engine noise. 
   Further, in this embodiment, as described above, the intake valve starts opening relatively late in this lower torque and speed condition, and generates the greater amplitude of the pressure wave as shown in  FIG. 14(A) . 
   In this embodiment, when the engine  1  operates in the engine idle condition with the desired engine torque below TQ D1  and the engine speed below N E2  as determined at the steps S 38  and S 40  of the routine R 3  shown in  FIG. 10 , the shutter valve  35  is opened at the step S 33 . The effective length of the pressure wave path is now the length between the intake valve  11  and the communication chamber  34 , and the shorter effective length causes more reflections and attenuations of the pressure wave before it reaches the intake port throat  9   a  when the intake valve opens. Therefore, the amplitude of the pressure wave may be reduced as shown by a solid line in  FIG. 14(A) . The inadvertent effect of the pressure wave, such as the cylinder air charge deviation. 
   It is needless to say that the invention is not limited to the illustrated embodiments and that various improvements and alternative designs are possible without departing from the substance of the invention as claimed in the attached claims. For example, the VVL mechanism of the intake valve drive mechanism  101  is not limited to the type illustrated mainly in  FIGS. 3 through 7  that has the peak valve lift retarded as the valve closing timing is retarded. But, it may be a type that has the peak valve lift timing fixed, and the valve opening and closing timings retarded and advanced symmetrically to each other. 
   The VCT mechanism  103  is not limited to the hydraulic actuator shown in  FIG. 3  and described above, but it may be an electric motor, an electromagnetic actuator, or any other pertinent in the art. Further, the intake valve drive mechanism  101  may have electromagnetic valve actuators for the individual intake valves  11 . 
   The communication chamber  34  is not limited to the one connects all of the branch intake passages  32  as shown in  FIG. 2 , but, for example, a volume chamber is provided for each of the branch intake passages  32 . Further, any other form of intake passage that can change the effective length of the pressure wave transmission can be used, for example one that changes a position of an opening of the branch intake passage  32  and  33  to the surge tank  31 . It can be realized by providing an opening the wall  36  between the communication chamber  34  and the surge tank  31 , or providing a telescopic structure at the end of the branch intake passage. 
   The characteristic of the opening of the shutter valve  35  at the lower desired air charge condition is not limited to that illustrated in  FIG. 15  where the shutter valve  35  is opened when the engine speed is less than the second predetermined speed N E2  and the desired torque is less than the predetermined torque TQ D1 , but only the engine speed less than N E2  may be used as the condition to open the shutter valve  35 .