Abstract:
Actuators and corresponding methods and systems for controlling such actuators offer efficient, fast, flexible control with large forces. In an exemplary embodiment, an fluid actuator includes a housing having first and second fluid ports, an actuation cylinder in the housing defining a longitudinal axis and having first and second ends in first and second directions, an actuation piston in the cylinder with first and second surfaces moveable along the longitudinal axis, a spring subsystem biasing the actuation piston to a neutral position, a first fluid space defined by the first end of the actuation cylinder and the first surface of the actuation piston, a second fluid space defined by the second end of the actuation cylinder and the second surface of the actuation piston; and a flow bypass that short-circuits the first and second fluid spaces when the actuation piston is not proximate to the second end of the actuation cylinder. A first flow mechanism controls fluid communication between the first fluid space and the first port, whereas a second flow mechanism controls fluid communication between the second fluid space and the second port. The first flow mechanism is always wide-open, whereas the second flow mechanism is open and closed when the flow bypass is closed and open, respectively. The system is able to latch the actuation piston at its second direction end position while making it possible for the actuation piston not to dwell at its first direction end position, thus reducing the overall actuation time.

Description:
REFERENCE TO RELATED APPLICATION 
   This application claims priority to Provisional U.S. Patent Application No. 60/809,117, file on May 26, 2006, the entire content of which are incorporated herein by reference. 

   FIELD OF THE INVENTION 
   This invention relates generally to actuators and corresponding methods and systems for controlling such actuators, and in particular, to actuators offering efficient, fast, flexible control with large forces. 
   BACKGROUND OF THE INVENTION 
   A split four-stroke cycle internal combustion engine is described in U.S. Pat. No. 6,543,225. It includes at least one power piston and a corresponding first or power cylinder, and at least one compression piston and a corresponding second or compression cylinder. The power piston reciprocates through a power stroke and an exhaust stroke of a four-stroke cycle, while the compression piston reciprocates through an intake stroke and a compression stroke. A pressure chamber or cross-over passage interconnects the compression and power cylinders, with an inlet check valve providing substantially one-way gas flow from the compression cylinder to the cross-over passage, and an outlet or cross-over valve providing gas flow communication between the cross-over passage and the power cylinder. The engine further includes an intake and an exhaust valve on the compression and power cylinders, respectively. The split-cycle engine according to the referenced patent and other related developments potentially offers many advantages in fuel efficiency, especially when integrated with an additional air storage tank interconnected with the cross-over passage, which makes it possible to operate the engine as an air hybrid engine. Relative to an electrical hybrid engine, an air hybrid engine can potentially offer as much, if not more, fuel economy benefits at much lower manufacturing and waste disposal costs. 
   To achieve the potential benefits, the air or air-fuel mixture in the cross-over passage has to be maintained at a predetermined firing condition pressure, e.g. approximately 270 psi or 18.6 bar gage-pressure, for the entire four stroke cycle. The pressure may go much higher to achieve better combustion efficiency. Also, the opening window of the cross-over valve has to be extremely narrow, especially at medium and high engine speeds. The cross-over valve opens when the power piston is at or near the top dead center (TDC) and closes very shortly after that. The total opening window in a split cycle engine may be as short as one to two milliseconds, compared with a minimum period of six to eight milliseconds in a conventional engine. To seal against a persistently high pressure in the cross-over passage, a practical cross-over valve is most likely a poppet or disk valve with an outward (i.e. away from the power cylinder, instead of into it) opening motion. When closed, the valve disk or head is pressured against the valve seat under the cross-over passage pressure. To open the valve, an actuator has to provide an extremely large opening force to overcome the pressure force on the head as well as the inertia. The pressure force will drop dramatically once the cross-over valve is open because of a substantial pressure-equalization between the cross-over passage and the power cylinder. Once the combustion is initiated, the valve should be closed as soon as desired to prevent the spread of the combustion into the cross-over passage, which also entails a need, during a certain period of combustion, to keep the valve seated against a power cylinder pressure that is higher than the cross-over passage pressure. In addition, the cross-over valve needs to be deactivated when the power stroke is not active in certain phases of the air hybrid operation. Like conventional engine valves, the seating velocity of the cross-over valve has to be kept under certain limit to reduce noise and maintain adequate durability. 
   In summary, the cross-over valve actuator has to offer a large opening force, a substantial seating force, a reasonable seating velocity, a high actuation speed, and timing flexibility while consuming minimum energy by itself Most, if not all, engine valve actuation systems are not able to meet these demands. 
   SUMMARY OF THE INVENTION 
   Briefly stated, in one aspect of the invention, one preferred embodiment of an fluid actuator includes a housing having first and second fluid ports, an actuation cylinder in the housing defining a longitudinal axis and having first and second ends in first and second directions, an actuation piston in the cylinder with first and second surfaces moveable along the longitudinal axis, a spring subsystem biasing the actuation piston to a neutral position, a first fluid space defined by the first end of the actuation cylinder and the first surface of the actuation piston, a second fluid space defined by the second end of the actuation cylinder and the second surface of the actuation piston; and a flow bypass that short-circuits the first and second fluid spaces when the actuation piston is not proximate to the second end of the actuation cylinder. A first flow mechanism controls fluid communication between the first fluid space and the first port, whereas a second flow mechanism controls fluid communication between the second fluid space and the second port. The first flow mechanism is always wide-open, whereas the second flow mechanism is open and closed when the flow bypass is closed and open, respectively. 
   In operation, the spring subsystem, the actuation piston, and the actuator load (e.g., an engine valve) work as a spring-mass pendulum system, efficiently converting, the potential energy in the spring subsystem to the kinetic energy in the moving mass and vice versa. The efficient energy conversion also leaves less energy for the snubbing mechanism to dissipate and provides better soft seating for the engine valve. The actuation efficiency is also greatly helped by the flow bypass, which, when effective, is able to minimize fluid flow and energy consumption. The system is able to latch the actuation piston at its second direction end position while making it possible for the actuation piston not to dwell at its first direction end position, thus reducing the overall actuation time. The actuator can be supplied and controlled by a 4-way actuation switch valve, two actuation 3-way valves, or one actuation 3-way valve. 
   In another embodiment, the second flow mechanism is always wide open, whereas the first flow mechanism is open and closed when the flow bypass is closed and open, respectively. The system is able to latch the actuation piston at its first direction end position while making it possible for the-piston not to dwell at its second direction end position. 
   In another embodiment, a spring controller allows the engine valve to close at power-off and provides a means for an effective start-up. 
   The present invention provides significant advantages over the prevailing fluid actuators and their control. Its ability to latch the actuator only at one end and allow a quick return motion at the other end greatly saves actuation time, which is important or critical in many applications, especially for the cross-over valve in an air hybrid engine. The fluid nature of the actuator provides high force and power density to deal with the demanding requirements, and yet the bypass mechanism is able to offer high energy efficiency. The control approaches associated with various switch valves are able to deal varying application needs, especially those for an air hybrid engine. With its pendulum arrangement, there is a centering or returning spring force available, in addition to a differential fluid force, to help open the engine valve. 
   The present invention, together with further objects and advantages, will be best understood by reference to the following detailed description taken in conjunction with the accompanying drawings. 

   
     BRIEF DESCRIPTION OF THE DRAWINGS 
       FIG. 1  is a schematic illustration of one preferred embodiment of the valve actuator, which is at its power-off state; 
       FIG. 2  is a schematic illustration of one preferred embodiment of the valve actuator, which is complete with initialization or in a closed state; 
       FIG. 3  is a schematic illustration of one preferred embodiment of the valve actuator, which is just opening up an engine valve, with the bypass being closed; 
       FIG. 4  is a schematic illustration of one preferred embodiment of the valve actuator, which has substantially opened an engine valve, with the bypass being open; 
       FIG. 5  is a schematic illustration of one preferred embodiment of the valve actuator, which is about to close an engine valve, with the bypass being closed; 
       FIG. 6  is a schematic illustration of another preferred embodiment which utilizes two actuation three-way valves; 
       FIG. 7  is a schematic illustration of another preferred embodiment which utilizes one actuation three-way valve and one optional startup switch valve; 
       FIG. 8  is a schematic illustration of another preferred embodiment which opens an engine valve in the second direction and utilizes one actuation three-way valve; 
       FIG. 9  is a schematic illustration of another preferred embodiment which opens an engine valve in the second direction and utilizes two actuation three-way valves; and 
       FIG. 10  is a schematic illustration of another preferred embodiment which opens an engine valve in the second direction and utilizes one actuation switch valve. 
   

   DETAILED DESCRIPTION OF THE INVENTION 
   Referring now to  FIG. 1 , a preferred embodiment of the invention provides an engine valve control system using one piston, one or more bypass passages, and a set of centering spring means. The system comprises an engine valve  20 , a fluid actuator  30 , an actuation switch valve  80 , a pair of actuation springs  71  and  72 , and a spring control  270 . 
   The actuation switch valve  80  supplies the fluid actuator  30  through a first port  61 , a first-port passage  104 , a second port  62 , and a second-port passage  106 . The first port  61  and the first-port passage  104  may be a physically or functionally continuous part, and so do the second port  62  and the second-port passage  106 . The valve  80  is a 2-position 4-way valve. It has four ports connected with a low-pressure P_L fluid line, a high-pressure P_H fluid line, the first-port passage  104 , and the second-port passage  106 . It is switched either to a left position  82  and a right position  84 . At the left position as shown in  FIG. 1 , the first-port and second-port passages  104  and  106  are in fluid communication with the P_H and P_L lines, respectively. At the right position (as shown in  FIG. 3 ), the first-port and second-port passages  104  and  106  are in fluid communication with the P_L and P_H lines, respectively. 
   The pressure P_H can be either constant or continuously variable. When variable, it is controlled to accommodate variability in system friction, engine valve opening, air pressure, the engine valve seating velocity requirement, etc. and/or to save operating energy when possible. A higher P_H value helps overcome higher system friction and air pressure force, and increase the engine valve opening, whereas a lower P_H value is better for softer seating of the engine valve and for saving energy. The pressure P_L can be simply the fluid tank pressure, the atmosphere pressure, or a fluid system backup pressure. The fluid system backup pressure can be supported or controlled, for example, by a spring-loaded check valve, with or without an accumulator. The P_L value is preferred to be as low as possible to increase the system efficiency, and yet high enough to help prevent fluid cavitation. 
   The engine valve  20  includes an engine valve head  22  and an engine valve stem  24 . The engine-valve head  22  includes a first surface  28  and a second surface  29 , which in the case of a split-cycle engine, are exposed to a cross-over passage  110  and the engine cylinder  102 , respectively. The engine valve  20  is operably connected with the fluid actuator  30  along a longitudinal axis  116  through the engine valve stem  24 , which is slideably disposed in an engine valve guide  120 . When the engine valve  20  is fully closed, the engine valve head  22  is in contact with an engine valve seat  26 , sealing off the fluid communication between the cross-over passage  110  and the engine cylinder  102 . 
   The fluid actuator  30  comprises an actuator housing  66 , within which, along the longitudinal axis  116  and from a first to a second direction (from the top to the bottom in the drawing), there are a first bore  44 , an actuation cylinder  52 , and a second bore  46 . The actuation cylinder  52  includes a bypass undercut  50 , a first end  56 , and a second end  54 . The second bore  46  is interrupted by a second-bore undercut  47 . Within these hollow elements from the first to the second direction lies a shaft assembly  31  comprising a first piston rod  34 , an actuation piston  32 , a second-piston-rod shoulder  38 , a second-piston-rod neck  40 , and a second piston rod  36 . The first and second piston rods  34  and  36  are slideably disposed in and substantially supported in the radial direction by the first and second bores  44  and  46 . The actuation piston  32  is slideably disposed in the actuation cylinder  52  when the shaft assembly  31  is at and near its second direction end of the stroke or travel. 
   The actuation piston  32  longitudinally divides the actuation cylinder  52  into a first fluid space  112  (between the actuation-cylinder first end  56  and the actuation-piston first surface  98 ) and a second fluid space  114  (between the actuation-piston second surface  100  and the actuation-cylinder second end  54 ). The two fluid spaces  112  and  114  are interconnected or in fluid communication through the bypass undercut  50  when the actuation-piston second surface  100  is over the bypass edge  58  in the first direction. When this flow bypass is effective, the two fluid spaces  112  and  114  and thus two actuation-piston surfaces  98  and  100  are under substantially the same pressure. This bypass effect does not substantially short-circuit the two ports  61  and  62  by making L 4  substantially equal to L 2 , where L 2  is the maximum longitudinal overlap between the actuation piston  32  and the non-bypass part of the actuation cylinder  52 , and L 4  is the maximum longitudinal underlap between the second piston rod  36  and the second bore  46 . 
   The radial clearances between the above sliding surfaces are substantially tight, provide substantial fluid seal, and yet offer tolerable resistance to relative motions, including translation along and, if desired, rotation around the longitudinal axis  116 , between the shaft assembly  31  and the housing  66 . 
   The actuation cylinder  52 , including the portion with the bypass undercut  50 , offers substantial axial length such that the actuation piston  32  does not contact its first end  56  at any operating condition. When the engine valve  20  is seated as shown in  FIG. 1 , there is still a distance L 3  between the actuation-piston second surface  100  and the actuation-cylinder second end  54  to accommodate the engine valve lash adjustment. 
   Concentrically wrapped around the engine valve stem  24  and the second piston rod  36 , respectively, are a first actuation spring  71  and a second actuation spring  72 . The second actuation spring  72  is supported by a housing surface  70  and a first spring retainer,  76 , whereas the first actuation spring  71  is supported by the first and a second spring retainers  76  and  78 . The actuation springs  71  and  72  are preferably under compression. 
   The first spring retainer  76  is operably connected with the engine valve stem  24  and the second piston rod  36 . Some part or element of this connection can be a simple mechanical contact as long as they move inseparably, which may be secured for example by designing proper spring preloads. If desired, the retainer  76  can be designed into two separate retainers as shown in  FIGS. 8-10 . 
   The second spring retainer  78  is supported by a spring controller  270 . The spring controller  270  includes a spring-controller bore  280  sliding over the engine valve guide  120  as shown in  FIG. 1 , or the engine valve stem  24  if the engine valve guide  120  is shorten toward the second direction as shown in  FIG. 6 . The controller  270  partitions a cavity in the engine cylinder head  68  into a spring-controller first and second chambers  272  and  274 . The second chamber  274  is supplied, through a spring-controller port  296 , with the working fluid from a fluid source P_SP. The second chamber  272  being preferably in communication with the atmosphere or a fluid return line (details of which not shown in  FIG. 1 ). Structurally, the spring controller  270  and its associated chambers  272  and  274  and port  296  can be alternatively supported by an extended part of the housing  66 , which is assembled on to the cylinder head  68 . 
   The longitudinal position of the spring controller  270  results primarily from the balance between the fluid pressure force on a spring-controller second surface  278  in the first direction and the spring force from the first actuation spring  71  in the second direction, and it is limited in the first and second directions when spring-controller first and second surfaces  276  and  278  come in contact with spring-controller chamber first and second surfaces  292  and  294  respectively. The pressure of the fluid source P_SP can be switched between a high value and a low value to position the spring controller  270  in two end positions in the first and second directions, respectively. If desired, the pressure of the fluid source P_SP can also be continuously controlled to situate the controller  270  in between its two end positions. If so, because of the variability of the spring force with the engine valve opening and closing, some damping mechanism (not shown in  FIG. 1 ) is needed to limit the position oscillation of the spring controller  270 . The fluid source S_SP can be simply the high pressure P_H line. Alternatively, it can tap into the engine lubrication supply system, and the same fluid is used to lubricate the engine valve stem  24  and the engine valve guide  120 . 
   When the spring controller  270  is at its second direction end position (as shown in  FIG. 1 ) because of a low or zero pressure in the second chamber  274  at a power-off state or during an actuator initialization, the two actuation springs  71  and  72  are at their least compressed state, and their static, net total force tends to move, by design, the engine valve  20  to a closed position, with an additional seating force if desired. When the spring controller  270  is at its first direction end position (as shown in  FIGS. 2-5 ) because of a high pressure in the second chamber  274 , the two springs  71  and  72  are together at their most compressed state, and their static, net total force tends to bias the engine valve  20 , in most designs, to a substantially middle point between the fully open and closed positions, setting up the system for its normal pendulum actuation. A position where the net or total spring force is zero is also called a neutral position. When desired, the engine valve neutral position can also be away from the substantial middle point between the fully open and closed positions. While the actuation springs  71  and  72  tend to bias the engine valve  20  to a neutral position, the actual position is also influenced by fluid forces on the actuation piston  32 , the air forces on the engine valve head  22 , inertia force during opening and closing, etc. The two springs  71  and  72  can be either identical or not in their designs and force curves. 
   The second-piston-rod shoulder  38  is intended to work with the second bore  46  as a snubber to slow down the shaft assembly  31  near the end of its travel in the second direction. When traveling in the first direction, the second actuation spring  72  always stalls the shaft assembly  31  and the engine valve  20  well before the actuation-piston first surface  98  is able to contact the actuation-cylinder first end  56 . 
   The shaft assembly  31  is generally under two longitudinal fluid forces on the actuation-piston first and second surfaces  98  and  100 . The effective pressure areas of the two surfaces  98  and  100  are influenced by the diameters of the first and second piston rods  34  and  36 . As an option, the actuator can be designed without the first piston rod to provide much more effective pressure area on the surface  98 . In  FIG. 1 , a first chamber  45 , distal to a first-piston-rod end surface  42 , is either in direct communication with the atmosphere, or with a fluid tank  108  through a third port  63  to collect the leaked fluid. The first-piston-rod, end surface  42  is therefore not exposed to any substantial pressure or pressure force. 
   The engine valve head  22  is generally exposed to the pressure of the crossover valve passage on the first surface  28  and the pressure of the engine cylinder  102  on the second surface  29 . 
   The system also experiences various friction forces, steady-state flow forces, transient flow forces, and other inertia forces. Steady-state flow forces are caused by the hydrostatic pressure redistribution due to flow induced velocity variation, i.e. the Bernoulli effect. Transient flow forces are fluid inertial forces. Other inertial forces result from the acceleration of objects, excluding fluid here, with inertia, and they are substantial in an engine valve assembly because of the large magnitude of the acceleration or the fast timing. 
   The fluid flow control within the actuator  30  can be considered to include a first flow mechanism, a second flow mechanism, and a bypass. The first and second flow mechanisms control fluid communication between the first fluid space  112  and the first port  61  and that between the second fluid space  114  and the second port  62 , respectively. The bypass controls fluid communication between the first and second fluid spaces  112  and  114 . For the preferred embodiment illustrated in  FIG. 1 , the first flow mechanism is direct (not shown in  FIG. 1 ) or indirect (through the bypass undercut  50 ) connection between the first fluid space  112  and the first port  61 . The second flow mechanism includes the second-bore undercut.  47 , the annular space between the second bore  46  and the second-piston-rod neck  40 , the second piston rod  36 , and the second-piston-rod shoulder  38 . The bypass includes the bypass undercut  50 . The first flow mechanism is always open, while the second flow mechanism is substantially open and closed when the bypass is closed and open, respectively. 
   Power-Off State 
   At power-off, all fluid supply sources, including the fluid supply source P_SP, are at low or zero gage pressure. The spring controller  270  is thus at the second direction end position as shown in  FIG. 1  under the spring force from the first actuation spring  71 . The total fluid force on the actuation piston  32  is substantially equal to zero. The actuation springs  71  and  72  are at their least compressed states, and their net force urges the engine valve  20  in a closed position as shown in  FIG. 1 , with the contact force from the engine valve seat  26  balancing the net spring force if assuming a negligible gravitational force. The level of the contact force can be pre-designed. When the actuation springs  71  and  72  are substantially symmetric or identical in their designs, the seating contact force is equal to the total spring stiffness times the spring length differential (Lsp 1 −Lsp 2 ). The spring lengths Lsp 1  and Lsp 2  should be substantially equal if the contact force approaches zero. 
   At the power-off, the actuation switch valve  80  is preferably, but, not necessarily, in its left position  82  as shown in  FIG. 1 , so that it does not have to be switched at the start-up. 
   Start-Up 
   To start-up the system from the power-off state as shown in  FIG. 1 , all fluid supply sources, including that of P_SP, are pressurized, the actuation switch valve  80  is secured at (or switched to if not already at the power-off state) its left position  82 , resulting in high and low pressures in the first and second fluid spaces  112  and  114 , respectively, and thus a net fluid force on the actuation piston  32  in the second direction. Also, the spring controller  270  is pushed by the fluid pressure in the spring-controller second chamber  274  to its first direction end position to increase the overall compression of the first and second actuation springs  71  and  72  and to move the neutral position of the pendulum in the first direction. As shown in  FIG. 2 , the length Lsp 1  of the first actuation spring  71  is now appreciably shorter than the length Lsp 2  of the second actuation spring  72  once the start-up is complete although the two springs  71  and  72  in this case are assumed to have identical design. 
   It is beneficial to establish the differential fluid force on the actuation piston  32  and latch the piston in its position as shown in  FIG. 2  before there is a substantial movement of the spring controller  270  in the first direction. This time delay may happen naturally since it takes much more fluid volume and thus more time to fill an expanding or moving volume like the spring-controller second chamber  274  than to pressurize a fixed volume like that of the actuation cylinder  52  and its bypass undercut  50 . This time delay can also be enforced by various means. For example, either another switch valve (not shown in  FIG. 2 ) or an artificially restrictive flow passage can be added between the spring-controller port  296  and the fluid source P_PS. For simplicity, one fluid source P_PS and one delay switch valve can be used by several actuators in an engine. 
   Valve Opening and Closing 
   To open the engine valve  20 , the actuation switch valve  80  is turned to the right position  84  as shown in  FIG. 3 , wherein the first and second ports  61  and  62  are connected with the low pressure P_L and high pressure P_H, respectively. Due to the open communication with the second port  62  through the second bore undercut  47  and the annular space between the second bore  46  and the second-piston-rod neck  40 , the pressure in the second fluid space  114  rises quickly from the low pressure P_L to a value close to the high pressure P_H. Due to open fluid communication with the first port  61 , the pressure in the first fluid space  112  drops quickly from the high pressure P_H to a value close to the low pressure P_L. The resulting net pressure force on the actuation piston  32  works with the net spring force to drive the shaft assembly  31  and the engine valve  20  in the first direction, overcoming the cross-over passage air pressure force on the engine valve first surface  28 . The net fluid pressure on the actuation piston  32  remains to be substantial and in the first direction as long as the engine valve displacement Xev remains to be less than L 2 . 
   When the engine valve displacement Xev exceeds L 2  as shown in  FIG. 4 , the second piston rod  36  completely overlaps the second-bore undercut  47  and blocks fluid communication between the second fluid space  114  and the second port  62 , resulting in a zero fluid flow through the second port  62 . At the same time, the actuation-piston second surface  100  is over the bypass edge  58  in the first direction, opening up the fluid communication between the two fluid spaces  112  and  114  and equalizing the pressure on the actuation-piston first and second surfaces  98  and  100 . The bypass flow between the two fluid spaces  112  and  114  substantially reduces the need for fluid flow to or from the first port  61 . When the first and second piston rods  34  and  36  have substantially the same diameter, there is substantially zero fluid flow through the first port  61 , and the net pressure force on the actuation piston  32  is substantially, equal to zero if the flow resistance in the bypass undercut  50  is ignored. At this state, the shaft assembly  31  and the engine valve  20  continue to travel in the first direction against the net spring force and potentially some residual differential air pressure force on the engine valve  20 , with their speed being reduced and most of their kinetic energy being converted into the potential energy in the springs  71  and  72 . 
   Eventually, the shaft assembly  31  and the engine valve  20  become completely stalled and reach their maximum displacement in the first direction (not shown in the figures). After that point, they start traveling back in the second direction under the spring force, converting the potential energy in the springs  71  and  72  back into the kinetic energy for the shaft assembly  31  and the engine valve  20 . 
   To get ready to latching the returning actuation piston  32  in the second direction end position, the actuation switch valve  80  is switched back into the left position  82 , preferably before the actuation-piston second surface  100  passes under the bypass edge  58  in the second direction, i.e., before the bypass is disabled. The displacement-time curve of the engine valve  20  is not substantially sensitive to the exact time of this switch action as long as it occurs while the bypass is still effective. 
   Once the bypass is disabled, the second piston rod  36  underlaps at least part of the second-bore undercut  47  as shown in  FIG. 5 , and the first and second fluid spaces  112  and  114  are in fluid communication with the high-pressure P_H and low-pressure P_L supply lines. The second-piston-rod shoulder  38  is intended to work with the second bore  46  as a snubbing mechanism to restrict the flow out of the second fluid space  114  and reduce the seating velocity of the engine valve  20  near the end of the stroke in the second direction. To further assist the snubbing effect, one may add restrictive features or size limits  86  and  88  to the left passages or ports in the actuation switch valve  80 . If the approaching speed of the actuation piston  32  is fast, the resulting larger pressure drops across the restrictive passages or ports  86  and  88  induce a smaller differential fluid pressure across the actuation piston  32  in the second direction. The differential fluid pressure force on the piston may point in the first direction when the approaching speed is too fast. The actuation springs  71  and  72  also work to slow down the engine valve  20 , with their net force being in the first direction during this part of the stroke, converting the kinetic energy of the shaft assembly  31  and the engine valve  20  into the spring potential energy. With the shoulder  38 , the springs  71  and  72 , and the restrictive ports  86  and  88 , the engine valve  20  is slowed down to an acceptable seating velocity. The restrictive features or ports  86  and  88  or other snubbing mechanism can also be designed to be thermally sensitive so that they are more restrictive at higher fluid temperature to compensate for temperature variation. 
   While the snubbing is in action, the actuation piston  32  is also prevented from backing away from seating and is eventually latched because the differential, pressure recovers substantially close to a value of (P_H−P_L) as soon as the piston  32  slows down. The resulting differential fluid pressure force on the actuation piston  32  is designed to be sufficient to counter the returning spring force in the first direction plus any differential air pressure force across the engine valve head  22 . The air pressure in the engine cylinder  102  may exceed the air pressure in the cross-over passage  110  in certain part of the combustion period, resulting in a net air pressure force in the first direction. After the combustion, the engine cylinder pressure drops rapidly while the cross-over passage pressure remains substantially high, when the net air pressure force helps keep the engine valve  20  seated. At the latched position, the state of the valve actuation system is back to the same state as that depicted in  FIG. 2 , ready for the next cycle. 
   In this invention, no attempt is made to hold or latch the actuation piston  32  and thus the engine valve  20  at their maximum opening position. They are driven substantially by the actuation springs  71  and  72  alone once the bypass is effective, and they are driven back in the second direction as soon as reaching their maximum opening, i.e., no dwell time. It is so designed to reduce the total opening and closing time, which is highly desired for the cross-over valve in an air hybrid engine. Without spending energy to latching the valve at the maximum opening, it also saves energy. 
     FIG. 6  depicts an alternative embodiment of the invention that utilizes first and second actuation 3-way valves  180  and  182 , instead of one 4-way switch valve  80  as in  FIGS. 1-5 . The first and second actuation 3-way valves  180  and  182  control the fluid communication to the first and second ports  61  and  62 , respectively. Using two separate valves provides more control flexibility. The two ports  61  and  62  do not necessarily have to have the opposite pressure polarities. For example, both first and second ports  61  and  62  may be at the low-pressure P_L (as shown in  FIG. 6 ), which reduces fluid leakage, when the engine valve  20  is securely seated under a high cross-over passage air pressure and a low engine cylinder air pressure. 
   The embodiment in  FIG. 6  also features a spring controller passage  298  that provides fluid communication between the cross-over passage  110  and the spring-controller second chamber  274 , which provides an alternative way to control the spring controller  270 . When the power being off and the cross-over passage  110  and thus the spring-controller second chamber  274  being out of pressurized gas or air, the spring controller  270  is situated at the second direction end position, resulting in a spring neutral position shifted in the second direction and a seated engine valve  20  under the spring forces. When the cross-over passage  110  being at a moderate to high pressure, the same pressure will be present in the spring-controller second chamber  274 , resulting in appropriately compressed actuation springs  71  and  72  as shown in  FIG. 6 , fit for the normal pendulum operation. 
   The embodiment in  FIG. 6  illustrates that the spring controller  270  can slide over the engine valve stem  24 , instead of the extended engine valve guide  120  as in  FIGS. 1-5 . 
     FIG. 7  depicts another preferred embodiment of the invention that features three major variations. First, the first port  61  is connected only with the high pressure P_H supply line, without control of a switch valve. During the opening stroke of the engine valve  20  in the first direction, both sides of the actuation piston  32  will be under the high pressure P_H, and the moving mass will be driven only by the actuation springs  71  and  72 , without the help of a favorable differential fluid force. This variation only needs one, instead of two, actuation  3 -way valve  182 , a much simpler arrangement. It is suitable for applications where frictional energy losses (including those to overcome the differential air force on the engine valve  20 ) are not substantial. The energy losses during the opening stroke (the travel in the first direction) can be compensated during the closing stroke (the travel in the second direction) when the flow bypass is not effective. 
   Second, the flow bypass in the actuation cylinder  52  is realized by at least one bypass passage  202 . The actuation cylinder  52  is longitudinally divided, by the actuation piston  32 , into first and second fluid spaces  112  and  114 . The function of the bypass undercut  50  (as shown in  FIG. 1 ) is replaced by the at least one bypass passage  202 , which opens up fluid communication between the first and second fluid spaces  112  and  114  when the actuation-piston second surface  100  moves over, in the first direction, a bypass edge  58 . The at least one bypass passage  202  is preferably to be geometrically axial-symmetric to result in axial-symmetric or balanced fluid forces on the shaft assembly  31 . If two bypass passages  202  are used, for example, they are preferably to be  180  degree apart. The first port  61  is connected with the first fluid space  112  either directly (not shown in  FIG. 7 ) or indirectly through at least one bypass passage  202  (as shown in  FIG. 7 ). 
   The third feature of the embodiment illustrated in  FIG. 7  is its lack of the spring controller  270  (as shown in  FIG. 1 ). The function of initializing the actuator, i.e. seating the engine valve  20  at the engine start-up, can be completed by enclosing the first bore  44  structurally and pressurizing with the first-piston-rod end surface  42  with a startup switch valve  192 . At normal operation, the startup switch valve  192  is kept at its right position to expose the first-piston-rod end surface  42  with the low pressure P_L to minimize energy losses. The same startup switch valve  192  can be shared by more than one actuator. As an alternative, the actuator  30  may be started without the startup switch valve  192 . Instead, the actuator  30  may be designed in such a way that the actuation-piston second surface  100  is still not over, in the first direction, the bypass edge  58  when the power is off. At the startup, the second fluid space  114  is not in fluid communication with the first fluid space  112 , and the high pressure P_H in the first fluid space  112  and the low pressure P_L in the second fluid space  114  are able to drive the shaft assembly  31  in the second direction and seat the engine valve  20 . 
   Refer now to  FIG. 8 , which is a drawing of yet another alternative embodiment of the invention. In this case, the engine valve  20   b  is opened in the second direction as in most conventional internal combustion engines. Therefore, the bypass undercut  50   b  is longitudinally located at the second direction end of the actuation cylinder  52   b.  The second fluid space  114   b  is in un-interrupted fluid communication with the second port  62   b  through the bypass undercut  50   b,  whereas the first fluid space  112   b  is in fluid communication with the first port  61   b  and the second fluid space  114   b,  respectively, when the actuation-piston first surface  98   b  is over (in the first direction) and under (in the second direction) the bypass edge  58   b,  i.e. when the bypass is ineffective and effective. The fluid communication between the first fluid space  112   b  and the first port  61   b  is through a first-bore undercut  48  and an annular space between the first bore  44   b  and the first-piston-rod neck  41 . A first-piston-rod shoulder  39  works with the first bore  44   b  to provide snubbing action needed for soft seating of the engine valve  20   b  in the first direction. The spring controller  270   b  is slideably disposed in the housing  66   b  and over the second piston rod  36   b  to control the position of the first direction end of the second actuation spring  72 . The spring controller  270   b  is longitudinally balanced by the force from the second actuation spring  72 , the fluid force from the spring-controller second chamber  274   b,  and a contact force when it is longitudinally limited by the housing  66   b.  The spring-controller second chamber  274   b  is supplied either from a separate fluid supply (not shown in  FIG. 8 ) or from the second port  62   b  through a spring-controller passage  297  and, optionally, the bypass undercut  50   b  or the second fluid space  114   b.  Like the embodiment illustrated in  FIG. 7 , one has the option of not using the spring controller at all. Instead, the startup may be achieved either by using a startup switch valve like the valve  192  shown in  FIG. 7  or simply keeping, by design, the bypass substantially ineffective when the net spring force is zero. 
   The first port  61   b  is connected to the high-pressure P_H and low-pressure P_L fluid lines through the second actuation 3-way valve  182   b,  whereas the second port  62   b  is directly connected with the high-pressure P_H fluid line. 
   For the preferred embodiment illustrated in  FIG. 8 , the second flow mechanism is a wide-open connection, through the bypass undercut  50   b,  between the second fluid space  114   b  and the second port  62   b.  The first flow mechanism includes the first-bore undercut  48 , the annular space between the first bore  44   b  and the first-piston-rod neck  41 , the first piston rod  34   b,  and the first-piston-rod shoulder  39 . The bypass includes the bypass undercut  50   b.  The second flow mechanism is always open, while the first flow mechanism is substantially open and closed when the bypass is closed and open, respectively. 
   At the engine power-off state, both fluid sources P_H and P_L are at or near zero gage pressure, and the spring-controller second chamber  274   b  is not pressurized. The spring controller  270   b  is therefore at its first direction end position (as shown in  FIG. 8 ), resulting in a closed engine valve as shown in  FIG. 8 . At the engine start-up, the second actuation 3-way valve  182   b  is at its right position (as shown in  FIG. 8 ), resulting in a differential pressure in the first direction on the actuation piston  32   b  to lock up the piston  32   b  in the first direction end position, resulting in a closed engine valve  20   b.  To open the engine valve  20   b,  the second actuation 3-way valve  182   b  is switched to its left position, thus feeding the high pressure P_H fluid to the first port  61   b  and the first fluid space  112   b  and substantially equalizing the pressure at the both sides of the actuation piston  32   b.  The piston  32   b  is then driven by the net spring force to move in the second direction, opening up the engine valve  20   b.  The net spring force decreases its magnitude as the moving mass gaining its speed. Once passing the neutral point where the net spring force is zero, the net spring force turns to be in the first direction and increases its magnitude as the moving mass loses its velocity until being stalled by the spring force. Then the spring-mass pendulum swings in the first direction until the actuation-piston first surface  98   b  passes, in the first direction, the bypass edge  58   b,  by then the second actuation 3-way valve  182   b  has been switched back to its right position to supply the first port  61   b  and the first fluid space  112   b  with the low pressure P_L fluid to create a differentials pressure force in the first direction, which helps keep the actuation piston  32   b  moving in the first direction. At the same time, the seating velocity is limited due to flow restriction created by the first-piston-rod shoulder  39  partially blocking the first bore  44   b  and optional flow restriction at the low pressure metering path in the second actuation 3-way valve  182   b.    
   The diameters of the first and second piston rods  34   b  and  36   b  do not have to be identical. It is preferable to have a relatively smaller diameter for the first piston rod  34   b  if more engine valve opening force is desired. 
   Refer now to  FIG. 9 , which is a drawing of yet another alternative embodiment of the invention. This embodiment in  FIG. 9  is different from that in  FIG. 8  primarily in the addition of a first actuation 3-way valve  180   b.  The first actuation 3-way valve  180   b  in  FIG. 9  is substantially the same as the first actuation 3-way valve  180  in  FIG. 6 . It offers the option of exposing the first and second fluid space  112   b  and  114   b  to the high-pressure P_H and low-pressure P_L, respectively, and thus having a differential pressure force in the second direction on the actuation piston  32   b  to help open the engine valve  20   b    
   Refer now to  FIG. 10 , which is a drawing of yet another alternative embodiment of the invention. This embodiment in  FIG. 10  is different from that in  FIG. 8  primarily in the use of an actuation switch valve  80   b,  instead of the second actuation 3-way valve  182   b.  The actuation switch valve  80   b  in  FIG. 10  is substantially the same as the actuation switch valve  80  in  FIGS. 1-5  in terms of physical structure and functions, which are explained in details earlier in this application. 
   In all the above descriptions, the first and second actuation springs  71  and  72  are each identified or illustrated, for convenience, as a single spring. When needed for strength, durability or packaging, however each or any one of the first and second actuation springs  71  and  72  may include a combination of two or more springs. In the case of mechanical compression springs, they can be nested concentrically, for example. The two actuation springs can also be combined into a single mechanical spring (not shown) that can take both tension and compression. They may also include a combination of pneumatic and mechanical springs, or even two pneumatic springs. The two springs can be either identical or not in their designs and force curves. The spring subsystem, either with a single or multiple springs, tends to return the shaft assembly to a neutral position. As a design option, the pneumatic springs may be filled, supplemented, or controlled by the pressurized air or gaseous mixture in the cross-over passage  110 . The pneumatic springs may have adjustable mass or pressure to achieve variable spring rate and thus variable valve stroke slope. Use of a pneumatic spring can also help close the engine valve  20  at power-off and startup the valve system. If the first actuation spring  71  in  FIG. 1  is a pneumatic one, for example, it can be discharged at power-off to bias the engine valve  20  in the second direction to a seated position, which also helps get the actuator ready for the next startup. After the next startup, the pneumatic spring will be charged again. 
   In all the above descriptions, each of the switch and/or control valves may be either a single-stage type or a multiple-stage type. Each valve can be either a linear type (such as a spool valve) or a rotary type. Each valve can be driven by an electric, electromagnetic, mechanic, piezoelectric, or fluid means. 
   In some illustrations and descriptions, the fluid medium may be assumed or implied to be in hydraulic or in liquid form. In most cases, the same concepts can be applied, with proper scaling, to pneumatic actuators and systems. As such, the term “fluid” as used herein is meant to include both liquids and gases. Also, in many illustrations and descriptions so far, the application of the invention is defaulted to be in engine valve control, and it is not limited so. The invention can be applied to other situations where a fast and/or energy efficient control of the motion is needed. 
   Although the present invention has been described with reference to the preferred embodiments, those skilled in the art will recognize that changes may be made in form and detail without departing from the spirit and scope of the invention. As such, it is intended that the foregoing detailed description be regarded as illustrative rather than limiting and that it is the appended claims, including all equivalents thereof, which are intended to define the scope of this invention.