Abstract:
An improved control method for a solenoid-operated hydraulic actuator for deactivating an engine valve mechanism characterizes the dynamic response of the mechanism based on a lumped parameter model of the solenoid, the hydraulic sub-system, and a locking pin mechanism actuated by a control pressure developed by the hydraulic sub-system. In response to a mode change request, constituent delay times associated with the solenoid, the hydraulic sub-system, and the locking pin mechanism are determined and summed to form an estimate of the overall delay time required to complete the requested cylinder deactivation. The solenoid activation is then scheduled based on the estimated delay time and a window of opportunity in the engine cycle for cylinder deactivation.

Description:
This application claims the benefit of Provisional Application No. 60/234,863, filed Sep. 22, 2000. 
    
    
     TECHNICAL FIELD 
     The present invention is directed to selective deactivation of specified cylinders of an internal combustion engine with a solenoid activated hydraulic actuator that disables intake and exhaust valve lifters, and more particularly to a model-based method of controlling the actuator based on an estimation of the actuator response time. 
     BACKGROUND OF THE INVENTION 
     It is known that fuel economy improvements may be achieved in multi-cylinder internal combustion engines by deactivating selected cylinders during specified engine operating conditions. For example, General Motors Corporation produced engines for 1980 Cadillac vehicles capable of operation with four, six or eight cylinders, depending on engine speed and load. In mechanizing a cylinder deactivation system in an engine with cam-driven valve lifters, the valve lifters for the intake and exhaust valves of a cylinder capable of being deactivated are equipped with solenoid activated hydraulic actuators that when activated, prevent the valves from opening. 
     In operation, the valve deactivation hardware is actuated in response to a command to deactivate the respective valves, and the actuation must be completed within a given window of opportunity relative to the respective cylinder combustion cycle. In a typical system, for example, the intake and exhaust valves for a cylinder to be deactivated are locked in a closed state during the combustion/power stroke, and prior to the exhaust stroke. To reliably carry out such a method, the controller must have a reasonably accurate estimation of the dynamic response time of the deactivation hardware. However, it is difficult to accurately calibrate or estimate the dynamic response time since it can vary significantly due to variations in the fluid source pressure, temperature, system voltage, and so on. Accordingly, what is needed is a control method based on a more accurate estimation of the overall response time of cylinder deactivation. 
     SUMMARY OF THE INVENTION 
     The present invention is directed to an improved control method in which the dynamic response of a solenoid-operated hydraulic actuator for deactivating an engine valve mechanism is characterized using a lumped parameter model of the solenoid, the hydraulic sub-system, and a locking pin mechanism actuated by a control pressure developed by the hydraulic sub-system. In response to a mode change request, constituent delay times associated with the solenoid, the hydraulic sub-system, and the locking pin mechanism are determined and summed to form an estimate of the overall delay time required to complete the requested cylinder deactivation. The solenoid activation is then scheduled based on the estimated delay time and a window of opportunity in the engine cycle for cylinder deactivation. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     FIG. 1A is a diagram of a prior art engine cylinder deactivation system including hydraulic intake and exhaust valve lifters equipped with solenoid operated hydraulic actuators for selectively preventing valve opening, and a microprocessor-based control unit for controlling solenoid energization in response to a cylinder activation/deactivation command. 
     FIG. 1B is a cross-sectional diagram of a hydraulic actuator of FIG.  1 A. 
     FIG. 2 is a block diagram of a dynamic model according to this invention. 
     FIG. 3, Graphs A-E, depict an example of cylinder deactivation according to this invention. 
     FIG. 4 is a flow diagram of a software routine executed by the control unit of FIG. 1 in carrying out a cylinder deactivation control according to this invention. 
    
    
     DESCRIPTION OF THE PREFERRED EMBODIMENT 
     Referring to FIG. 1A-1B, the method of the present invention is described in the context of a valvetrain deactivation system  10  in which intake and exhaust valves of specified engine cylinders are locked in a closed position to essentially deactivate the specified cylinders. In general, cylinder selection for deactivation is determined by engine firing order and the desire to maintain an even firing order after the specified cylinders have been deactivated. In the illustrated embodiment, the engine being controlled has two banks of four cylinders (i.e., a V8 engine); two specified cylinders from each bank are subject to selective deactivation, and the specified cylinders are consecutively deactivated in response to a mode change request. Also, the illustrated engine has only one intake valve and one exhaust valve per cylinder. 
     FIG. 1A depicts the intake and exhaust valve lifters IL 1 , EL 1 ; IL 4 , EL 4 ; IL 6 , EL 6 ; and IL 7 , EL 7  for cylinder numbers  1 ,  4 ,  6  and  7  of a V 8  engine, each such lifter being equipped with a hydraulically activated locking pin mechanism  12 , as depicted in FIG.  1 B. Each such valve lifter is mechanically actuated by a respective lobe  18   a ,  18   b ;  18   c ,  18   d ;  20   a ,  20   b ;  20   c ,  20   d  of a rotating camshaft, and in each case, the locking pin mechanism  12  can be positioned to either make or break a mechanical connection between the respective camshaft lobe and a respective valve. Referring to FIG. 1B, the locking pin mechanism  12  may be configured as a pair of shoes  22 ,  24  slidably disposed within a cavity  26  of an outer case  28  of an intake or exhaust lifter. The outer case  28  is mechanically coupled to a valve, and an inner member  30  slidably disposed within the axial bore of outer case  28  is mechanically coupled to a respective camshaft lobe. The shoes  22 ,  24  are biased apart by a spring  32  to a position that permits the transfer of linear motion from the camshaft lobe to the valve through the inner member  30 , the shoes  22 ,  24  and the outer case  28 . Hydraulic pressure supplied to a port  34  of the cavity  26  acts on the peripheral surfaces of shoes  22 ,  24 , producing a force that opposes spring  32 ; if the inner member  30  is in the depicted position (that is, if the base circle of the camshaft lobe is contacting the inner member  30 ), the shoes  22 ,  24  are free to move, and the hydraulic force displaces the shoes  22 ,  24  inwardly until limited by the stops  36 ,  38 . In this position, the shoes  22 ,  24  no longer couple inner member  30  to outer case  28 ; consequently, linear motion of the camshaft lobe is no longer coupled to the respective valve, and such valve remains closed. Instead, the shoes  22 ,  24  transfer the axial motion of inner member  30  to inner member  40 , which is biased to the position shown in FIG. 1B to maintain the shoes  22 ,  24  properly positioned within the cavity  26  when the inner members  30  and  40  move back to the depicted positions. When the camshaft lobe is contacting a valve lifter on a point other than the base circle of the lobe, the shoes  22 ,  24  are transmitting cam motion as described above, and cannot be deactivated. 
     Referring back to FIG. 1A, the valvetrain deactivation system  10  additionally comprises a hydraulic system  50  capable of coupling a source of fluid pressure Ps (which may be engine oil pressure) to the locking pin mechanism  12  of each lifter IL 1 , EL 1 ; IL 4 , EL 4 ; IL 6 , EL 6 ; IL 7 , EL 7 , a set of solenoid-operated valves  52 ,  54 ,  56 ,  58 , and a microprocessor-based engine control module (ECM)  60  for activating the solenoid-operated valves  52 ,  54 ,  56 ,  58  in response to a mode change request. The fluid pressure Ps is supplied to a supply pressure plenum  62 , and fluid in the plenum  62  is supplied to control chambers  64 ,  66 ,  68 ,  70  for each of the cylinders subject to deactivation through respective charge orifices  72 ,  74 ,  76 ,  78 . The solenoid-operated valves  52 ,  54 ,  56 ,  58  are schematically depicted, each including an electrically-activated solenoid coil S 1 , S 4 , S 6 , S 7  electrically coupled to ECM  60  and a shiftable fluid conduit  80 ,  82 ,  84 ,  86  for selectively coupling the respective control chamber  64 ,  66 ,  68 ,  70  to an exhaust orifice  88 ,  90 ,  92 ,  94  or to a supply pressure orifice  96 ,  98 ,  100 ,  102 . And in each case, the shiftable conduit  80 ,  82 ,  84 ,  86  is mechanically biased to a position that exhausts the respective control chamber  64 ,  66 ,  68 ,  70 . The fluid pressure in each control chamber  64 ,  66 ,  68 ,  70  thus has a base or default value determined by the relative areas of charge orifices  72 ,  74 ,  76 ,  78  and exhaust orifices  88 ,  90 ,  92 ,  94  that is insufficient to overcome the force of spring  32  in the respective locking pin mechanism  12 . However, when one or more of the solenoid coils S 1 , S 4 , S 6 , S 7  are electrically activated to change the valve state, the fluid pressure in the respective control chamber  64 ,  66 ,  68 ,  70  increases substantially to the supply pressure Ps, producing a force that is sufficient to overcome the force of spring  32  in the respective locking pin mechanism  12 , provided that the inner member  30  is contacting the base circle of the respective camshaft lobe. 
     When controlling an engine equipped with the above-described system  10 , the overall time response of the system must be known in order to ensure that the specified cylinders are reliably activated and deactivated, and in order to coordinate the deactivation hardware with other engine control functions such as spark timing and fuel delivery. For example, when a change mode request is generated, ECM  60  must determine the overall response time of the system  10  in order to determine if there is time to activate the mechanical pin mechanisms  12  in the up-coming window of opportunity in the engine crank cycle. If there is not sufficient time, the ECM  60  must wait until the next window of opportunity. If there is sufficient time, the ECM  60  must command activation of the solenoid valves  52 ,  54 ,  56 ,  58 , disable fuel delivery to the specified cylinders, and suitably adjust the spark timing for the other cylinders. Unfortunately, accurate characterization of the overall response time is not easily achieved since laboratory testing cannot encompass the various combinations of engine and system parameters that affect response time. According to the present invention, this difficulty is overcome by modeling the dynamic behavior of the deactivation system hardware to accurately estimate the overall response time for any combination of the various parameters that affect response time. The overall response time includes three components: the solenoid valve response time, the hydraulic system response time, and the locking pin mechanism response time. 
     As shown in FIG. 2, the system  10  can essentially be modeled as a series of three subsystems: solenoid valve subsystem  110 , a hydraulic subsystem model  112 , and locking pin subsystem  114 . On receipt of a mode change request (i.e., a request to deactivate specified engine cylinders), ECM  60  supplies a solenoid activation voltage signal v to the solenoid valve subsystem  110 , which outputs a valve spool displacement signal x sp  indicative of the resulting movement of a solenoid valve  52 ,  54 ,  56 ,  58 . The output x sp  is applied as an input to the hydraulic subsystem model  112  along with other inputs indicative of the air ratio A/R in control chambers  64 ,  66 ,  68 ,  70 , the supply pressure P S  and the fluid temperature T oil . Based on such inputs, the hydraulic system model  112  outputs a control pressure signal Pc indicative of the resulting pressure in a control chamber  64 ,  66 ,  68 ,  70 . The Pc signal is supplied as an input to the locking pin subsystem model  114 , which outputs a displacement signal Xp indicative of the resulting movement of shoes  22 ,  24 . The overall response time begins when mode change request is received and ends when the signal Xp indicates that the shoes  22 ,  24  have been fully displaced. 
     The solenoid subsystem model  110  takes into account the mechanical, electrical, and electromagnetic aspects of the solenoid valves  52 ,  54 ,  56 ,  58 . The mechanical aspects are represented by the force balance equation: 
     
       
           m   sp   ·x   sp   +B   sp   ·x   sp   +K   sp   ·x   sp   =F   EMF   −F   f   −F   preload   (1) 
       
     
     where: 
     x sp  is the solenoid plunger displacement 
     m sp  is the mass of solenoid plunger 
     B sp  is the damping coefficient of solenoid plunger 
     K sp  is the spring coefficient of solenoid plunger spring 
     F EMF  is the electromagnetic force 
     F f  is the fluid force 
     F preload  is the solenoid spring preload 
     The electrical aspects of the solenoid valves are represented by the equation:              v   =       R   ·     i        (   t   )         +       L        (   x   )              ∂   i       ∂   t         +       i        (   t   )                ∂     L        (   x   )           ∂   x       ·          x          t                     (   2   )                                
     where R is the solenoid coil resistance, i is the solenoid coil current, L is the solenoid coil inductance, and v is the solenoid coil excitation voltage. 
     Finally, the electromagnetic force F EMF  is given by the equation:                F   EMF     =       2        μ   0        π                   N   2          i   2           (         4        x   sp       d     +       l   g       d   +     l   g           )     2               (   3   )                                
     where μ 0  is the air permeability, d is the diameter of solenoid plunger, l g  is the air gap, and N is the number of solenoid coil turns. 
     Since the design parameters of the solenoid valves  52 ,  54 ,  56 ,  58  are known, the dynamic response of the solenoid valves can be characterized by equations (1), (2) and (3). Certain of these parameters, such as the solenoid coil resistance R, change with temperature, and the coil temperature may be approximated by fluid temperature T oil . The solenoid response time Δt sol  is the time it takes for the plunger displacement x sp  to reach a predetermined fully actuated displacement. Ignoring the fluid force F f , ECM  60  can characterize the solenoid plunger response time as a function of the solenoid voltage λ and the fluid temperature T oil . That is: 
     
       
         Δ t   sol   =f   1 ( ν,T   oil )  (4) 
       
     
     In the hydraulic model  112 , the supply pressure P s  is an input, and the flow continuity equation for control chambers  64 ,  66 ,  68 ,  70  can be written as:                  Q     sc                 1       +     Q     sc                 2       -     Q   leak       =                V   1          (   t   )              t       +           V   1          (   t   )         β   e       ·            P   1            t                   (   5   )                                              Q     sc                 1       =       C   q     ·     a     sc                 1       ·         2        (       P     s                 1       -     P     c                 1         )       ρ                 (   6   )                                               Q     sc                 2       =       C   q     ·     a     sc                 2       ·         2        (       P     s                 1       -     P     c                 1         )       ρ       ·     u        (     t   -     t     sol                 1         )                 (   7   )                                 V   1   =V   10   +a   pin   ·x   pin     —     in   ·u ( t−t   lin )+ a   pin   ·x   pin     —     ex   ·u ( t−t   lex )  (8) 
     where P sl  is the supply pressure at the control chamber, assuming equal to P s , u is a step function, i.e.                u        (     t   -     t     sol                 1         )       =     {         0         t   &lt;     t     sol                 1                 1         t   ≥     t     sol                 1                         (   9   )                                
     t sol1  is the time instant when the solenoid valve is energized, t lin  is the time instant when the respective intake lifter sits on the cam base circle, and is a function of crank timing, t lex  is the time instant when the respective exhaust lifter sits on the camshaft base circle, and is a function of crank timing, and β e  be is the equivalent bulk modulus of the engine oil. Assuming the air in the engine oil is homogeneous, the equivalent bulk modulus can be calculated by:                1     β   e       =       1     β   f       +     υ        1     β   g                   (   10   )                                
     where β f  is the bulk modulus of the engine oil, β g  is the adiabatic bulk modulus, which is 1.4 P for air. υ is the air ratio, i.e. the ratio of air volume to the total volume. Additionally, Q leak  is the leakage flow, and is given by:                Q   leak     =             (     C   rad     )     3     ·   Δ                     P   ·     d   m     ·   π         12          μ   m     ·   l                 (   11   )                                
     where C rad  is the radial clearance,          d   m     l                          
     is the ratio of clearance&#39;s mean diameter to the land length, and μ m  is the mean absolute viscosity. 
     Assuming the plenum  62  has uniform pressure P s , then equation (8) can be modified as follows to take into account the pressure fluctuation as a result of intake and exhaust lifter locking pins moving at different time instants.                V   1     =       V   10     +       a   pm     ·       ∑       i   =   1     ,   4   ,   6   ,   7              x   ipin_in     ·     u        (     t   -     t     i                 n         )             +       a   pin     ·       ∑       j   =   1     ,   4   ,   6   ,   7              x   ipin_ex     ·     u        (     t   -     t   jex       )                       (   12   )                                
     The hydraulic system response time Δt h  is simply the time for the control pressure to rise to a critical level determined by design criteria, and can be characterized as a function of the supply pressure P s  and the fluid temperature T oil . That is: 
     
       
         Δ t   h   =f   2 ( P   s   ,T   oil )  (13) 
       
     
     The locking pin model considers the pin mechanism  12  as a mass-spring-damper system described by the dynamic equations:              {                 m   pin     ·       x   ¨     1       +       B   pin     ·       x   .     1       +       K   pin     ·     (       x   1     +     x   2       )         =     F   1                       m   pin     ·       x   ¨     2       +       B   pin     ·       x   .     2       +       K   pin     ·     (       x   1     +     x   2       )         =     F   2                     (   14   )                                
     where x 1  and x 2  are the displacements of shoes  22  and  24 , and forces F 1  and F 2  are the hydraulic forces applied to shoes  22  and  24 . The transfer functions relating x 1  and x 2  to the input forces F 1  and F 2  are as follows:                  x   1          (   s   )       =           (         m   pin          s   2       +       B   pin        s     +     K   pin       )     ·       F   1          (   s   )         -       K   pin     ·       F   2          (   s   )                 m   pin   2          s   4       +     2        B   pin          m   pin          s   3       +       (       B   pin   2     +     2        K   pin          m   pin         )          s   2       +     2        K   pin          B   pin                   (   15   )                   x   2          (   s   )       =           (         m   pin          s   2       +       B   pin        s     +     K   pin       )     ·       F   2          (   s   )         -       K   pin     ·       F   1          (   s   )                 m   pin   2          s   4       +     2        B   pin          m   pin          s   3       +       (       B   pin   2     +     2        K   pin          m   pin         )          s   2       +     2        K   pin          B   pin                   (   16   )                 F   1     =         P   c     ·     a   pin       -     F   preload               (   17   )                 F   2     =       (         P   c     ·     a   pin       -     F   preload       )     ·     u        (     t   -   T     )                 (   18   )                                
     where F preload  is the pin spring preload, T is the time delay between pressure forces F 1  and F 2 , and s is the Laplace operator. If T is small enough, then e −Ts ≈1. Then the pin motion equations (15) and (16) are simplified as:                  x   1          (   s   )       =       -       x   2          (   s   )         =       1         m   pin          s   2       +       B   pin        s     +     2        K   pin           ·       F   1          (   s   )                   (   19   )                                
     Equation (17) indicates that if the fluid enters the left hand side of cavity  26  fast enough that the fluid force can be considered to act on both pins simultaneously, the two-shoe spring system can be modeled by a simplified one shoe system with twice the equivalent spring constant. The locking pin response time Δt p  is the time it takes for the shoes  22 ,  24  to reach the stops  36 ,  38  once the control pressure rises to the critical pressure level, and can be characterized by ECM  60  as a function of control pressure Pc and fluid pressure T oil . That is: 
     
       
         Δt p   =f   3 ( P   c   ,T   oil )  (20) 
       
     
     The total response time of the deactivation hardware system is then defined as: 
     
       
           Q=Δt   sol   +Δt   h   +Δt   p   (21) 
       
     
     Finally, a fourth response time to consider is the ECM response time Δ t   s , which may be negligible compared to the other response times. 
     In summary, ECM  60  estimates the overall response time for each of the specified engine cylinders as the sum of four constituent response times as graphically depicted in FIG.  3 . Referring to FIG. 3, a mode change request occurs at time t 0 , causing ECM  60  to determine the overall response time, and to determine if there is sufficient time to complete the requested cylinder deactivation by the end of the next engine cycle window of opportunity. In the example of FIG. 3, there is sufficient time, and a command to energize the solenoid S 1  is issued at time t 1 , resulting in complete activation of the corresponding locking pin mechanism  12  at time t 5 . Similarly, commands for energizing the solenoids S 4 , S 6  and S 7  are issued at times t 2 , t 3  and t 4 , respectively, resulting in complete activation of the respective locking pin mechanisms at times t 6 , t 7  and t 8 . 
     A flow diagram representative of a software routine periodically executed by ECM  60  for carrying out the above-described control is depicted in FIG.  4 . Referring to FIG. 4, the block  100  initially determines if a mode change request (that is, a signal requesting deactivation of specified engine cylinders) has occurred. If not, the routine is exited; if so, the block  102  is executed to determine the overall delay times ODT 1 , ODT 7 , ODT 6  and ODT 4  for cylinder numbers  1 ,  7 ,  6  and  4 . In practice, the overall delay times for each cylinder may be assumed to be equal for any given set of environmental and engine operating conditions. The delay times may be determined in a straight-forward manner based on equations 1-3, 5-12, 14 and 17-19 above, or may be determined by table-look-up based on T oil , V, Ps and Pc as indicated in equations  4 ,  13  and  20 . In the latter case, the table values are determined by solving the equations 1-3, 5-12, 14 and 17-19 off-line for various combinations of T oil , V and Ps, as will be well understood by those skilled in the art. The block  104  then determines if there is sufficient time to complete the deactivation of cylinder number  1  in the next crank window of opportunity; that is, whether the difference between the end of the crank window (CRANK_WINDOW 1 ) and the current crank angle (CRANK_CURRENT) is greater than the overall delay time ODT  1 . If not, the routine is exited, and cylinder deactivation is delayed until the following window of opportunity. If so, the blocks  106 ,  108 ,  110  and  112  are executed to schedule the respective solenoid activation times based on the determined delay times and the respective crank windows, and the block  114  is executed to clear the mode change request, and to set flags indicating that the cylinder deactivation commands have been issued. 
     While the present invention has been described in reference to the illustrated embodiments, it is expected that various modifications in addition to those mentioned above will occur to those skilled in the art. For example, the described control is applicable to other types of engines and other control strategies including a bank control in which the specified engine cylinders are concurrently deactivated instead of consecutively deactivated. Thus, it will be understood that control method incorporating these and other modifications may fall within the scope of this invention, which is defined by the appended claims.