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, and a second fluid space defined by the second end of the actuation cylinder and the second surface of the actuation piston. 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 and second flow mechanisms are substantially restricted through two integrated snubbing mechanisms when the actuation piston approaches the first and second direction ends of its travel, respectively. In addition to a differential fluid force on the actuation piston, there is a centering or returning spring force available to help open the engine valve against the high cross-over passage pressure, without the need for the fluid actuation system to be bulky and consume too much energy.

Description:
REFERENCE TO RELATED APPLICATION 
   This application claims priority to Provisional U.S. Patent Application No. 60/841,038, file on Aug. 30, 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 and U.S. Publication No. US2005/0016475A1. 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 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 a 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, and a second fluid space defined by the second end of the actuation cylinder and the second surface of the actuation piston. 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 and second flow mechanisms are substantially restricted through two integrated snubbing mechanisms when the actuation piston approaches the first and second direction ends of its travel, 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 mechanisms to dissipate and provides better soft seating for the engine valve. The actuation efficiency is further helped by utilizing two actuation 3-way valves, with one of them being switched to the high pressure fluid purposely at a later time during the engine valve return travel. 
   The system is able to latch the actuation piston at each end of its travel. The actuation piston does not have to, if desired, contact the end of the actuation cylinder for it to be latched. The piston may achieve a substanially steady balance simply through a combination of fluid forces and the net spring force. 
   In another embodiment, the actuator is supplied and controlled by a 4-way actuation switch valve. Each of the 4-way and 3-way valves may be a proportional valve when desired. 
   In another embodiment, a spring controller allows the engine valve to close at power-off even without sufficient pressure in the cross-over passage. 
   The present invention provides significant advantages over the prevailing fluid actuators and their control. Its ability to latch the actuator at both ends is important or critical in applications where an engine valve has to be held at open for a controllable period of time. The fluid nature of the actuator provides high force and power density to deal with the demanding requirements of a cross-over valve, and yet the spring-pendulum mechanism is able to offer high energy efficiency. The control approaches associated with various switch valves are able to deal with 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 against the high cross-over passage pressure, without the need for the fluid actuation system to be bulky and consume too much energy. 
   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 a closed state; 
       FIG. 2  is a schematic illustration of one preferred embodiment of the valve actuator, which starts opening up an engine valve; 
       FIG. 3  is a schematic illustration of one preferred embodiment of the valve actuator, which starts closing an engine valve; 
       FIG. 4  is a schematic illustration of another preferred embodiment, which utilizes one four-way actuation valve and Belleville springs, and offers a variation in flow mechanism design; 
       FIG. 5  is a schematic illustration of another preferred embodiment, which includes two piston rods with different diameters; 
       FIG. 6  is a schematic illustration of another preferred embodiment, which utilizes a proportional valve for control; and 
       FIG. 7  is a schematic illustration of another preferred embodiment which opens an engine valve in the second direction. 
   

   DETAILED DESCRIPTION OF THE INVENTION 
   Referring now to  FIG. 1 , a preferred embodiment of the invention provides an engine valve control system using one actuation piston, and a set of centering spring means. The system comprises an engine valve  20 , a fluid actuator  30 , a first actuation 3-way valve  180 , a second actuation 3-way valve  182 , a pair of actuation springs  71  and  72 . 
   The first and second actuation 3-way valves  180  and  182  supply the fluid actuator  30  through a first port  61  (via a first-port passage  104 ) and a second port  62  (via a second-port passage  106 ), respectively. 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 . Each of the 3-way valves  180  and  182  has two ports connected with a low-pressure P_L fluid line and a high-pressure P_H fluid line, and the third or remaining port connected with one of the two port passages  104  and  106 . 
   The 3-way valve  180  is switched either to a left position  184  or a right position  186 . At the left and right positions  184  and  186 , the first-port  61  is in fluid communication with the P_H and P_L lines, respectively. The 3-way valve  182  is switched either to a left position  188  or a right position  190 . At the left and right positions  188  and  190 , the second-port  62  is in fluid communication with the P_H and P_L 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 speed, whereas a lower P_H value is better for softer seating of the engine valve and for saving energy. The low 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 simply 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 or starvation. When necessary, the low pressure P_L can be more tightly controlled as well. 
   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 first end  56  and a second end  54 . The first and second bores  44  and  46  are interrupted by a first-bore undercut  48  and a second-bore undercut  47 , respectively. Within these hollow elements from the first to the second direction lies a shaft assembly  31  comprising a first piston rod  34 , a first-piston-rod neck  41 , a first-piston-rod shoulder  39 , 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 , respectively. The actuation piston  32  is slideably disposed in the actuation cylinder  52 . 
   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 piston  32  includes a first surface  98  and a second surface  100 , and longitudinally divides the actuation cylinder  52  into a first fluid space  112  (a fluid volume between the actuation-cylinder first end  56  and the actuation-piston first surface  98 ) and a second fluid space  114  (a fluid volume between the actuation-piston second surface  100  and the actuation-cylinder second end  54 ). 
   The fluid actuator  30  further includes a first reed valve  200  and a second reed valve  202 . The first reed valve  200  provides substantially one-way fluid communication from the first port  61  to the first fluid space  112 , which is facilitated by an actuation-cylinder first undercut  58 . The second reed valve  202  provides substantially one-way fluid communication from the second port  62  to the second fluid space  114 , which is facilitated by an actuation-cylinder second undercut  60 . 
   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 the housing  66  (or any spring retaining feature, not shown in  FIG. 1 , connected with the housing  66 ) and a central spring retainer  76 , whereas the first actuation spring  71  is supported by the central spring retainer  76 , and the cylinder head  68  (or any spring retaining feature, not shown in  FIG. 1 , connected with the cylinder head  68 ). The actuation springs  71  and  72  are preferably under compression. 
   The central 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 (not shown in the figures). 
   The first-piston-rod and second-piston-rod shoulders  39  and  38  are intended to work with the first and second bores  44  and  46  as snubbing or flow-restricting mechanism to slow down the shaft assembly  31  near the end of its travel in the first and second directions, respectively. 
   The actuation cylinder  52  offers substantial room in the second direction such that the actuation piston  32  does not contact its second end  54  at any operating condition. When the engine valve  20  is seated as shown in  FIG. 1 , there is still a longitudinal distance between the actuation-piston second surface  100  and the actuation-cylinder second end  54  to accommodate the engine valve lash adjustment. 
   In the first direction, there are two design and operating options. In the first option, the shaft assembly  31  is balanced at the steady state by fluid forces and the net spring force before the actuation-piston first surface  98  reaches the actuation-cylinder first end  56 . In the second option, the shaft assembly  31  is balanced at the steady state by fluid forces, the net spring force, and the contact force resulting from the contact between the actuation-piston first surface  98  and 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 . A first chamber  45 , distal to a first-piston-rod end surface  42 , is either in communication with a fluid tank  108  through a third port  63  to collect the leaked fluid as shown in  FIG. 1 , or in direct communication with the atmosphere (see  FIG. 4 ). The fluid tank  108  is preferably the same tank the rest of the fluid system uses. 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 the first and second reed valves  200  and  202 . The first flow mechanism and the first reed valve  200  control fluid communication between the first fluid space  112  and the first port  61 . The second flow mechanism and the second reed valve  2002  control fluid communication between the second fluid space  114  and the second port  62 . 
   The first flow mechanism, for the embodiment illustrated in  FIG. 1 , involves the first-bore undercut  48 , an annular space between the first bore  44  and the first-piston-rod neck  41 , the first piston rod  34 , and the first-piston-rod shoulder  39 . The first flow mechanism is substantially open when the annular space between the first bore  44  and the first-piston-rod neck  41  is substantially open both to the first fluid space  112  and the first-bore undercut  48 . When the actuation piston  32  is near or at the first direction end of its travel, the first-piston-rod shoulder  39  protrudes into the annular space between the first bore  44  and the first-piston-rod neck  41 , resulting in flow restriction and thus snubbing function. The underlap X 12  between the first-bore undercut  48  and the first piston rod  34  is generally of a sufficient length, regardless the position of the piston  32  so as not to cause flow restriction. If necessary or desired, the underlap X 12  can be designed to be substantially short when the actuation piston  32  is near or at the second direction end of its travel (as shown in  FIG. 1 ) to introduce a certain amount of flow restriction. The first reed valve  200  is optional and is intended to allow for a one-way flow from the first port  61  to the first fluid space  112  to bypass the flow restriction through the first flow mechanism, helping quickly fill the first fluid space  112  at the beginning of the piston travel in the second direction. The first flow mechanism may optionally not include the first-bore undercut  48 , with the first-piston-rod neck  41  being extended further in the first direction so that the annular space between the first bore  44  and the first-piston-rod neck  41  directly opens to the first port  61 . 
   The second flow mechanism, for the embodiment illustrated in  FIG. 1 , involves the second-bore undercut  47 , an 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 second flow mechanism is substantially open when the annular space between the second bore  46  and the second-piston-rod neck  40  is substantially open both to the second fluid space  114  and the second-bore undercut  47 . When the actuation piston  32  is near or at the second direction end of its travel as shown in  FIG. 1 , the second-piston-rod shoulder  38  protrudes into the annular space between the second bore  46  and the second-piston-rod neck  40 , resulting in flow restriction and thus snubbing function. The underlap X 22  between the second-bore undercut  47  and the second piston rod  36  is generally of a sufficient length, regardless the position of the piston  32  so as not to cause flow restriction. If necessary or desired, the underlap X 22  can be designed to be substantially short when the actuation piston  32  is near or at the first direction end of its travel to introduce a certain amount of flow restriction. The second reed valve  202  is optional and is intended to allow for a one-way flow from the second port  62  to the second fluid space  114  to bypass the flow restriction through the second flow mechanism, helping quickly fill the second fluid space  114  at the beginning of the piston travel in the first direction. The second flow mechanism may optionally not include the second-bore undercut  47 , with the second-piston-rod neck  40  being extended further in the second direction so that the annular space between the second bore  46  and the second-piston-rod neck  40  directly opens to the second port  62 . 
   Power-Off State 
   There are two possible power-off states for the fluid actuator  30  in a split cycle engine. One of them is when the engine or power is off while the cross-over passage  110  is still sufficiently pressurized, especially for an air-hybrid application with an air storage tank. The high and low pressure fluid sources P_H and P_L are all at low or zero gage pressure. The total fluid force on the actuation piston  32  is substantially equal to zero. Still, the pressure in the cross-over passage  110  is able to overcome the centering spring force, hold the engine valve  20  against the valve seat  26 , and keep the fluid actuator  30  in a state substantially like that shown in  FIG. 1 . 
   At the other power-off state, when the cross-over passage  110  is not sufficiently pressurized, the engine valve is balanced primarily by the net spring force and stays about half open (not shown in  FIG. 1 ). The actuation piston  32  is half-way between its two end positions. 
   At the power-off, the first and second actuation 3-way valves  180  and  182  are preferably, but not necessarily, in their left and right positions  184  and  190 , respectively, as shown in  FIG. 1  so that they do not have to be switched at the next start-up. 
   Start-Up 
   To start-up the system from the power-off state, all fluid supply sources are pressurized, and the actuation 3-way valves  180  and  182  are secured at their positions as shown in  FIG. 1 , which then leads a differential pressure between the first and second fluid spaces  112  and  114 , causing the engine valve  20  either to be secured at or to be driven to a closed position as shown in  FIG. 1 . 
   Valve Opening and Closing 
   To open the engine valve  20 , the first and second actuation 3-way valves  180  and  182  are switched to their right and left positions  186  and  188 , respectively, as shown in  FIG. 2 . The first fluid space  112  is in communication with the low pressure P_L supply through the first flow mechanism. The first reed valve is kept closed because of an unfavorable pressure direction. The second fluid space  114  is in communication with the high pressure P_H supply through the second flow mechanism and the second reed valve  202 , which is under a differential pressure in favor of opening and helps alleviate potential cavitation or starvation in the second fluid space  114 , especially during the initial period of the travel when the second flow mechanism is restrictive. The differential pressure force on the actuation piston  32  works with the net spring returning force in the first direction to overcome the differential air pressure force on the engine valve, which is in the second direction because of the high pressure in the cross-over passage  110 . 
   The actuation piston  32  travels from the second-direction end position to its first-direction end position, the net spring force changes from its maximum return force in the first direction to its maximum return force in the second direction. The net spring force can be zero either at the central point of the travel or, if desired, at a point which is off the center. In air-hybrid engine applications, the pressure in the cross-over passage  110  is substantially constant because of the air storage tank. The pressure in the engine cylinder  102  is initially low, increases rapidly as soon as the engine valve  20  opens, and eventually reaches a value substantially equal to the pressure in the cross-over passage  110 . 
   As the actuation piston  32  approaches its first-direction end position, the first-piston-rod shoulder  39  starts approaching or protruding into the first bore, increasing the flow resistance in the first flow mechanism, and causing a substantial pressure rise in the first fluid space  112 , resulting in a snubbing action to dramatically slow down the piston velocity. In addition, with the two-spring pendulum design, the speed of the shaft assembly  31  is already substantially reduced at this point due to an increasing net spring return force in the second direction. Finally, the system reaches a steady state, with the differential pressure force in the first direction balances out the net spring return force in the second direction, a much reduced differential air force on the engine valve, and potentially a contact force between the actuation-cylinder first end  56  and the actuation-piston first surface  98  if they are in contact either by design and/or by operating conditions. 
   The closing process of the engine valve  20  is substantially the opposite of the opening process. There are important differences though. Once the engine valve  20  is wide-open, there is not substantial pressure differential on the engine valve. The fluid actuator  30  does not have to overcome major air pressure force to close the engine valve  20 . To reduce the energy consumption and to help achieve softer engine valve seating or landing, one may optionally keep, during a substantial, initial period of the closing process, the first actuation 3-way valve  180  at its right position while switching the second actuation 3-way valve  182  to its right position as shown in  FIG. 3 , resulting in a substantially low differential fluid pressure on the actuation piston  32 . The closing motion is therefore substantially driven by the net spring return force alone during this initial period. The first actuation 3-way valve  180  can be switched to its left position  184  at a later stage or time of the engine valve closing process to secure and latch the engine valve  20  at the closed position, against the net spring return force in the first direction and the differential air pressure force on the engine valve  20  in the first direction, which happens when the engine cylinder pressure exceeds the cross-over passage pressure due to combustion. The exact timing for switching the first actuation 3-way valve  180  to its left position can be controlled based on engine operating conditions, including the engine RPM, load, and fluid temperature or viscosity. 
   The second-piston-rod shoulder  38  works with the second bore  46  to increase flow resistance in the second flow mechanism and to create a snubbing action during the engine valve seating process. 
     FIG. 4  depicts an alternative embodiment of the invention that utilizes one 4-way switch valve  80 , instead of the first and second actuation 3-way valves  180  and  182  as in  FIGS. 1-3 . The valve  80  is a 2-position 4-way valve. It has four ports connected with the low-pressure P_L fluid supply, the high-pressure P_H fluid supply, the first-port passage  104 , and the second-port passage  106 . It is switched either to a left position  82  or and a right position  84 . At the left position as shown in  FIG. 4 , 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 (not 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 embodiment in  FIG. 4  is equipped with the first and second actuation springs of the Belleville type  71   b  and  72   b , each of which includes at least one coned disk. In each spring, two or more coned disks may be stacked in series (as shown in  FIG. 4 ) or in parallel. 
   The embodiment in  FIG. 4  also features a spring controller  270 . The spring controller  270  includes a spring-controller bore  280  sliding over the engine valve stem  24  as shown in  FIG. 4 , or the engine valve guide  120  if the engine valve guide  120  is longitudinally extended in the first direction. The spring 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 first chamber  272  being preferably in communication with the atmosphere or a fluid return line (details of which not shown in  FIG. 4 ). 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   b  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. 4 ) 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. 4 ) 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   b  and  72   b  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 or contact force if desired. When the spring controller  270  is at its first-direction end position (not shown in  FIG. 4 ) because of a high pressure in the second chamber  274 , the two springs  71   b  and  72   b  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   b  and  72   b  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   b  and  72   b  can be either identical or not identical in their designs and force curves. 
   The embodiment in  FIG. 4  highlights the optional differential between the sizes or diameters of the first and second piston rods  34   b  and  36   b , with the first piston rod  34   b  being visibly larger than the second piston rod  36   b , resulting in an appreciably larger effective area on the actuation-piston second surface  100   b  than on the actuation-piston first surface  98   b , and thus higher differential or net fluid force in the first direction than in the second direction under the identical pressure differential. If desired, the design can be reversed with the first piston rod  34   b  being smaller than the second piston rod  36   b  (not shown in  FIG. 4 ) to achieve the opposite force effect. When desirable, one may completely eliminate the first piston rod  34   b  (not shown in  FIG. 4 ) to achieve a greater net fluid force in the second direction. 
   The embodiment in  FIG. 4  further features variations in the first and second flow mechanisms. The first-bore and second-bore undercuts  48   b  and  47   b  are extended longitudinally to the actuation-cylinder first and second ends  56  and  54 , respectively. With this extension, the first-piston-rod and second-piston-rod necks  41  and  40  featured in  FIGS. 1-3  are no longer necessary in  FIG. 4  for the purpose of fluid communication. For flow restriction, the first-piston-rod and second-piston-rod shoulders  39  and  38  now work with the first-bore and second-bore undercuts  48   b  and  47   b , respectively, instead of the first and second bores  44  and  46  as in  FIGS. 1-3 . 
   The embodiment in  FIG. 4  also shows variations in the one-way fluid communication means or check valves, which are designed as the first and second reed valves  200  and  202  in  FIGS. 1-3 . They are optional. The fluid actuator may include only one check valve  202   b  as shown in  FIG. 4  or no check valve at all. A check valve can be in the form of a reed valve shown in  FIGS. 1-3  or other designs, such as a spring-loaded ball valve  202   b  in  FIG. 4 . 
     FIG. 5  depicts an alternative embodiment of the invention that 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 as shown in  FIG. 5 , resulting in 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   b  and  72   b , fit for the normal pendulum operation. 
   Refer now to  FIG. 6 , which is a drawing of yet another alternative embodiment of the invention. In this fluid actuator  30   c , the first and second ports  61   c  and  62   c  are in direct fluid communication, respectively, with the actuation-cylinder first and second undercuts  58   c  and  60   c , which are situated longitudinally a short distance away from the actuation-cylinder first and second ends  56   c  and  54   c , respectively. 
   When the actuation-piston first surface  98   c  passes in the first direction the actuation-cylinder first undercut  58   c , it substantially traps a certain amount of fluid in the first fluid space  112   c  and the first bore undercut  48   c  and creates snubbing action. The extent of the snubbing action can be designed into a taper  50  on the actuation piston  32   c , which regulates the extent of flow leak back into the cylinder. The first bore undercut  48   c  is optional and is intended to work with an optional first check valve  200   c  to avoid cavitation or starvation when the actuation piston  32   c  moves away from the actuation-cylinder first end  56   c.    
   Similarly, when the actuation-piston second surface  100   c  passes in the second direction the actuation-cylinder second undercut  60   c , it substantially traps a certain amount of fluid in the second fluid space  114   c  and the second bore undercut  47   c  and creates snubbing action. The extent of the snubbing action can be designed into one or more slots  51  on the actuation piston  32   c , which regulates the extent of flow leak back into the cylinder. The slots  51  can also be placed on a wall of the actuation cylinder, instead of the piston. Also, the taper  50  and the slots  51  can be interchanged to achieve the same snubbing function. The second bore undercut  47   c  is optional and is intended to work with an optional second check valve  202   c  to avoid cavitation when the actuation piston  32   c  moves away from the actuation-cylinder second end  54   c . The first and second check valves  200   c  and  202   c  can be reed valves as shown in  FIG. 6 . 
   The embodiment in  FIG. 6  also features an actuation proportional valve  81 , which controls continuously the cross-section areas of its metering ports to achieve more controllability per performance requirements and operating conditions. While this proportional valve  81  is a 4-way valve, the actuation 3-way valves  180  and  182  featured in  FIGS. 1-3  may also be replaced with corresponding 3-way proportional valves. 
   Refer now to  FIG. 7 , which is a drawing of yet another alternative embodiment of the invention. In this case, the engine valve  20   d  is opened in the second direction as in most conventional internal combustion engines. When the engine valve  20   d  is closed as shown in  FIG. 7 , the actuation-piston first surface  98  is approximate to the actuation-cylinder first end  56 , and there is a gap between them for the engine valve lash adjustment. Most of variations of the invention discussed above and implied otherwise also apply to the embodiment in  FIG. 7 . 
   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 identical 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 start-up 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 start-up, the pneumatic spring will be charged again. Also when desired, one can physically separate the two actuation springs and place one of them, for example the second actuation spring  72  or  72   b  at the first direction end of the fluid actuator, where it can be operably connected with the first piston rod  34  or  34   b  or  34   c.    
   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.