Abstract:
Improved actuators and valve control systems, and methods for controlling actuators and/or engine valves, are disclosed. In addition to the inherent capability of timing control, the ability to provide continuous valve lift or stroke control greatly improves engine achieve fuel economy, emission and performance. The power-off state of the actuator is at the minimum stroke, from which an easy start-up can be directly executed. The minimum stroke is also very beneficial to achieve efficient low load operation. Even with continuous lift variation, the present invention is able to keep the spring force neutral or zero point in the center of a stroke, thus maintaining an efficient scheme of energy conversion and recovery through the pendulum action. When in compression braking or other high engine cylinder air pressure working mode, the invention is able to supply necessary force to open the engine valve. By adding a substantial hydraulic force to coincide with the spring returning force at the beginning of each stroke, the system can help overcome the engine cylinder air pressure and compensate for frictional losses. The invention incorporates lash adjustment into all alternative preferred embodiments, and makes it possible to trigger and complete one engine valve stroke by just one, instead of two, switch actions of the actuation switch valve.

Description:
REFERENCE TO RELATED APPLICATION 
   This application is a continuation-in-part of U.S. patent application Ser. No. 11/194,243, filed Aug. 1, 2005, the entire content of which is 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 providing independent lift (or stroke) and timing control with minimum energy consumption. 
   BACKGROUND OF THE INVENTION 
   Various systems can be used to actively control the lift (or stroke) and timing of engine valves to achieve improvements in engine performance, fuel economy, emissions, and other characteristics. Depending on the means of the control or the actuator, these systems can be classified as mechanical, electrohydraulic, and electromechanical (sometimes called electromagnetic). Depending on the extent of the control, they can be classified as variable valve-lift and timing, variable valve-timing, and variable valve-lift. They can also be classified as cam-based or indirect acting and camless or direct acting. 
   In the case of a cam-based system, the traditional engine cam system is kept and modified somewhat to indirectly adjust valve timing and/or lift. In a camless system, the traditional engine cam system is completely replaced with electrohydraulic or electromechanical actuators that directly drive individual engine valves. All current production variable valve systems are cam-based, although camless systems will offer broader controllability, such as cylinder and valve deactivation, and thus better fuel economy. 
   Problems with an electromechanical camless system include difficulty associated with soft-landing, high electrical power demand, inability or difficulty to control lift (or stroke), and limited ability to deal with high and/or varying cylinder air pressure. An electrohydraulic camless system can generally overcome such problems, but it does have its own problems such as performance at high engine speeds and design or control complexity, resulting from the conflict between the response time and flow capability. To operate at up to 6,000 to 7,000 rpm, an actuator has to first accelerate and then decelerate an engine valve over a range of 8 mm within a period of 2.5 to 3 milliseconds. The engine valve has to travel at a peak speed of about 5 m/s. These requirements have stretched the limit of conventional electrohydraulic technologies. 
   One way to overcome this performance limit is to incorporate, in an electrohydraulic system like in an electromechanical system, a pair of opposing springs which work with the moving mass of the system to create a spring-mass resonance or pendulum system. In the quiescent state, the opposing springs center an engine valve between its end positions, i.e., the open and closed positions. To keep the engine valve at one end position, the system has to have some latch mechanism to fight the net returning force from the spring pair, which accumulates potential energy at either of the two ends. When traveling from one end position to the other, the engine valve is first driven and accelerated by the spring returning force, powered by the spring-stored potential energy, until the mid of the stroke where it reaches its maximum speed and possesses the associated kinetic energy; and it then keeps moving forward fighting against the spring returning force, powered by the kinetic energy, until the other end, where its speed drops to zero, and the associated kinetic energy is converted to the spring-stored potential energy. 
   With its well known working principle, this spring-mass system by itself is very efficient in energy conversion and reliable. Much of the technical development has been to design an effective and reliable latch-release mechanism which can hold the engine valve to its open or closed position, release it as desired, add additional energy to compensate for frictions and highly variable engine cylinder air pressure, and damp out extra energy before its landing on the other end. As discussed above, there have been difficulties associated with electromechanical or electromagnetic latch-release devices. There has also been effort in the development of electrohydraulic latch-release devices. 
   Disclosed in U.S. Pat. No. 4,930,464, assigned to DaimlerChrysler, is an electrohydraulic actuator including a double-ended rod cylinder, a pair of opposing springs that tends to center the piston in the middle of the cylinder, and a bypass that short-circuits the two chambers of the cylinder over a large portion of the stroke where the hydraulic cylinder does not waste energy. When the engine valve is at the closed position, the bypass is not in effect, the piston divides the cylinder into a larger open-side chamber and a smaller closed-side chamber, and the engine valve can be latched when the open-side and closed-side chambers are exposed to high and low pressure sources, respectively, because of the resulting differential pressure force on the piston in opposite to the returning spring force. When the engine valve is at the open position, the piston divides the cylinder into a larger closed-side chamber and a smaller open-side chamber, and the engine valve can be latched by exposing a larger closed-side chamber and smaller open-side chamber with high and low pressure sources, respectively. 
   At either open or closed position, the engine valve is unlatched by briefly opening a 2-way trigger valve to release the pressure in the larger chamber and thus eliminate the differential pressure force on the piston, triggering the pendulum dynamics of the spring-mass system. The 2-way valve has to be closed very quickly again, before the stroke is over, so that the larger chamber pressure can be raised soon enough to latch the piston and thus the engine valve at its new end position. This configuration also has a 2-way boost valve to introduce extra driving force on the top end surface of the valve stem during the opening stroke. 
   The system just described has several potential problems. The 2-way trigger valve has to be opened and closed in a timely manner within a very short time period, no more than 3 ms. The 2-way boost valve is driven by differential pressure inside the two cylinder chambers, or stroke spaces as the inventers refer as, and there is potentially too much time delay and hydraulic transient waves between the boost valve and cylinder chambers. Near the end of each stroke, the larger cylinder chamber has to be back-filled by the fluid fed through a restrictor, which demands a fairly decent opening size on the part of the restrictor. On the other hand, at the onset of the each stroke, the 2-way trigger valve has to relieve the larger chamber which is in fluid communication with the high pressure fluid source through the same restrictor. During a closing stroke, there is no effective means to add additional hydraulic energy until near the very end of the stroke, which may be a problem if there are too much frictional losses. Also, this invention does not have means to adjust its lift. 
   DaimlerChrysler has also been assigned U.S. Pat. Nos. 5,595,148, 5,765,515, 5,809,950, 6,167,853, 6,491,007, and 6,601,552, which disclose improvements to the teachings of U.S. Pat. No. 4,930,464. The subject matter up to U.S. Pat. No. 6,167,853 resulted in various hydraulic spring means to add additional hydraulic energy at the beginning of the opening stroke to overcome engine cylinder air pressure force. One drawback of the hydraulic spring is its rapid pressure drop once the engine valve movement starts. 
   In U.S. Pat. No. 6,601,552, a pressure control means is provided to maintain a constant pressure in the hydraulic spring means over a variable portion of the valve lift, which however demands that the switch valve be turned between two positions within a very short period time, say 1 millisecond. The system again contains two compression springs: a first and second springs tend to drive the engine valve assembly to the closed and open positions, respectively. The hydraulic spring means is physically in serial with the second compression spring. During a substantial portion of an opening stroke, it is attempted to maintain the pressure in the hydraulic spring despite of the valve movement and thus provide additional driving force to overcome the engine cylinder air pressure and other friction, resulting in a net fluid volume increase in the hydraulic spring means and an effective preload increase in the second compression spring because of a force balance between the hydraulic and compression springs. In the following valve closing stroke, the engine valve may not be pushed all the way to a full closing because of higher resistance from the second compression spring. 
   A concern common to this entire family of inventions is that there have to be two switchover actions of the control valve for each opening or closing stroke. Another common issue is the length of the actuator with the two compression springs separated by a hydraulic spring. When the springs are aligned on the same axis, as disclosed in U.S. Pat. No. 5,809,950, the total height may be excessive. In the remaining patents of this family, the springs are not aligned on a straight axis, but are instead bent at the hydraulic spring, and the fluid inertia, frictional losses, and transient hydraulic waves and delays may become serious problems. Another common problem is that the closing stroke is driven by the spring pendulum energy only, and an existence of substantial frictional losses may pose a serious threat to the normal operation. As to the unlatching or release mechanism, some embodiments use a 3-way trigger valve to briefly pressurize the smaller chamber of the cylinder to equalize the pressure on both surfaces of the piston and reduce the differential pressure force on the piston from a favorable latching force to zero. Still the trigger valve has to perform two actions within a very short period of time. 
   U.S. Pat. No. 5,248,123 discloses another electrohydraulic actuator including a double-ended rod cylinder, a pair of opposing springs that tends to center the piston in the middle of the cylinder, and a bypass that short-circuits the two chambers of the cylinder over a large portion of the stroke where the hydraulic cylinder does not waste energy. Much like the referenced DaimlerChrysler patents, it has the larger chamber of the hydraulic cylinder connected to the high pressure supply all the time. Different from DaimlerChrysler, however, it uses a 5-way 2-position valve to initiate the valve switch and requires only one valve action per stroke. The valve has five external hydraulic lines: a low-pressure source line, a high-pressure source line, a constant high-pressure output line, and two other output lines that have opposite and switchable pressure values. The constant high pressure output line is connected with the larger chamber of the cylinder. The two other output lines are connected to the two ends of the cylinder and are selectively in communication with the smaller chamber of the cylinder. Much like the DaimlerChrysler disclosures, it has no effective means to add hydraulic energy at the beginning of a stroke to compensate for the engine cylinder air force and friction losses. It is not capable of adjusting valve lift either. 
   The actuators, and corresponding methods and systems for controlling such actuators described in my co-pending U.S. patent application Ser. No. 11/194,243, the entire content of which is incorporated herein by reference, provide independent lift and timing control with minimum energy consumption. In an exemplary embodiment, an actuation cylinder in a housing defines 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, is moveable along the longitudinal axis. First and second actuation springs bias the actuation piston in the first and second directions, respectively. A first fluid space is defined by the first end of the actuation cylinder and the first surface of the actuation piston, and a second fluid space is defined by the second end of the actuation cylinder and the second surface of the actuation piston. A fluid bypass short-circuits the first and second fluid spaces when the actuation piston is not substantially proximate to either the first or second end of the actuation cylinder. A first flow mechanism is provided in fluid communication between the first fluid space and a first port, and a second flow mechanism is provided in fluid communication between the second fluid space and a second port. The actuator may be coupled to a stem to form a variable valve actuator in an internal combustion engine, for example. 
   SUMMARY OF THE INVENTION 
   The present invention provides significant advantages over other actuators and valve control systems, and methods for controlling actuators and/or engine valves. In addition to the inherent capability of timing control, the ability of various embodiments to provide continuous valve lift or stroke control enhances engine fuel economy, emission and overall functionality. 
   By virtue of the invention, the power-off state of the actuator is at the minimum stroke, from which an easy start-up can be directly executed. The minimum stroke is also very beneficial to achieve efficient low load operation. Even with continuous lift variation, the present invention is able to keep the spring force neutral or zero point in the center of a stroke, thus maintaining an efficient scheme of energy conversion and recovery through the pendulum action. 
   By adding a substantial hydraulic force to coincide with the spring returning force at the beginning of each stroke, the system can help overcome the engine cylinder air pressure and compensate for frictional losses. The present invention is able to incorporate lash adjustment into all alternative preferred embodiments. It is also possible to trigger and complete one engine valve stroke by just one, instead of two, switch actions of the actuation switch valve. 
   One preferred embodiment of an electrohydraulic actuator according to the invention comprises a housing having first and second fluid ports, a stroke controller slideably disposed in the housing, first and second partial cylinders in the housing and the stroke controller, respectively, defining a longitudinal axis and having cylinder first and second ends in first and second directions, respectively, an actuation piston between the first and second partial cylinders with first and second surfaces moveable along the longitudinal axis, first and second actuation springs biasing the actuation piston in the first and second directions, respectively. 
   The actuator further includes a first fluid space defined by the cylinder first end and the piston first surface, a second fluid space defined by the cylinder second end and the piston second surface, a fluid bypass that short-circuits the first and second fluid spaces when the actuation piston does not overlap either of the first and second partial cylinders. Attached to the piston first surface are a first neck and a first piston rod, and attached to the piston second surface are a second neck and a second piston rod. The housing contains a first bore distal, in the first direction, to and in fluid communication with the first fluid space, whereas the stroke controller contains a second bore distal, in the second direction, to and in fluid communication with the second fluid space. A first chamber inside the housing is in fluid communication with the first port and the first bore, and a second chamber inside the stroke controller is in fluid communication with the second bore. A first groove is one or more undercuts situated between and in fluid communication with the second chamber and the second port and, independent of the longitudinal location of the stroke controller. 
   Traversing the first and second piston rods, respectively, are first and second rod passages which are in fluid communication with the fluid bypass via one or more center passages longitudinally inside the first and second piston rods, the first and second necks and the actuation piston and one or more piston passages traversing the actuation piston. A second-supplemental chamber is one or more undercuts around the first bore further distal, in the first direction, to the first chamber and in fluid communication with the second port, and a first supplemental chamber is one or more undercuts around the second bore, further distal, in the second direction, to the second chamber. A second groove is one or more undercuts situated between and in fluid communication with the first-supplemental chamber and the first port, independent of the longitudinal location of the stroke controller. 
   A first flow mechanism includes the first neck, the first piston rod, the first bore, and the first chamber, whereby controlling fluid communication between the first fluid space and the first port. A second flow mechanism includes the second neck, the second piston rod, the second bore, and the second chamber, whereby controlling fluid communication between the second fluid space and the second port. A first-supplemental flow mechanism includes the second groove, the first-supplemental chamber, the second rod passage, the center passage, the piston passage and the fluid bypass, whereby controlling fluid communication between the first fluid space and the first port. A second-supplemental flow mechanism includes the second-supplemental chamber, the first rod passage, the center passage, the piston passage and the fluid bypass, whereby controlling fluid communication between the second fluid space and the second port. 
   The actuator further comprises one or more snubbers, whereby the speed of the actuation piston is substantially damped when the piston travels approaching either of the cylinder first and second ends. An engine valve is operably connected to the second piston rod. 
   The inside dimension of the first bore is slightly larger than the outside dimension of the first piston rod and substantially larger than the outside dimension of the first neck, and the first piston rod blocks fluid communication between the first bore and the first chamber and thus closes the first flow mechanism when the actuation piston does not overlaps the first partial cylinder. The inside dimension of the second control bore is slightly larger than the outside dimension of the second rod and substantially larger than the outside dimension of the second neck, and the second piston rod blocks fluid communication between the second bore and the second chamber and thus closes the second flow mechanism, when the actuation piston does not overlaps the second partial cylinder. 
   The first-supplemental flow mechanism is opened when the second rod passage at least partially overlaps the first-supplemental chamber, which happens when the actuation piston overlaps the second partial cylinder; and the second-supplemental flow mechanism is opened when the first rod passage at least partially overlaps the second-supplemental chamber, which happens when the actuation piston overlaps the first partial cylinder. 
   The actuation piston can be latched to the cylinder first end, such that with the engine valve in a closed position, when the second and first fluid spaces are exposed to high- and low-pressure fluid, respectively, and not short-circuited by the fluid bypass because the resulting differential pressure force on the piston is in opposite to and greater than a returning force from the first and second actuation spring. Likewise, the actuation piston can be latched to the cylinder second end, such that with the engine valve in an open position, when the first and second fluid spaces are exposed to high- and low-pressure fluid, respectively, and not short-circuited by the bypass means. 
   At either open or closed position, the engine valve is unlatched or released by toggling an actuation switch valve so that the pressure levels in the first and second fluid spaces are reversed, instead of being equalized as in the prior art, and thus the differential pressure force on the piston is also reversed, instead of just being reduced to almost zero like in prior art. Before the switch, the differential pressure force on the actuation piston is in opposite to and greater than the spring returning force to latch the engine valve. After the switch, the differential pressure force keeps substantially the same magnitude and reverses its direction to help the spring returning force drive the engine valve to the other position, feeding additional hydraulic energy into the system. 
   By virtue of the invention, the position of the stroke controller and thus the stroke are controlled by a stroke spring and the pressure force in a stroke control chamber, in addition to the forces from the actuation springs and fluid pressure in the fluid bypass and the second fluid space. In alternative embodiments, they are directly controlled by mechanical means such as a set of rack and pinion or a set of mechanically driven pins. 
   In the embodiment described above, the first-supplemental and second-supplemental flow mechanisms comprise the passages along the axis of the first and second piston rods and through the actuation piston. In alternative embodiments, they only include passages through the stroke controller and the housing. 
   First and second shoulders situated between the necks and the piston end surfaces may be used to penetrate the first and second bores to restrict fluid communication and thus to create snubbing effect. Alternatively, a fluid trapping design at the first directional end of a capped first bore is used to offer substantial hydraulic force on the first directional end of the first piston rod before the engine valve lands on the valve seat. This additional snubbing action may also be switched on and off or controlled continuously by an optional end flow control mechanism, resulting in a varying degree of engine valve soft-landing required under different engine operating conditions. In another preferred embodiment, it is possible to selectively supply a high pressure to a fourth port connected to the piston first rod first end to provide additional driving force in the first direction. In yet another preferred embodiment, it is possible to design the two actuation springs with different preloads and/or spring rates to meet various functional needs, such as a closed engine valve at the power-off state or the net spring force biased more in the second direction to counter the biased engine cylinder air pressure force. In still another preferred embodiment, the first-supplemental and second-supplemental flow mechanisms are implemented with a 3-way shuttle valve, resulting in a more compact design. 
   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 one hydraulic actuator and hydraulic supply system according to the invention; 
       FIG. 2   a  is a schematic illustration of a hydraulic actuator with a first flow mechanism and second supplemental flow mechanism being open when an actuation piston overlaps with a first partial cylinder; 
       FIG. 2   b  is a schematic illustration of a hydraulic actuator with a second flow mechanism and first supplemental flow mechanism being open when an actuation piston overlaps with a second partial cylinder; 
       FIG. 3  is a schematic illustration of one preferred embodiment of the hydraulic actuator, which is complete with initialization. The engine valve is in closed position; 
       FIG. 4  is a schematic illustration of one preferred embodiment of the hydraulic actuator, with the maximum stroke and at the beginning of an opening stroke or travel in the second direction; 
       FIG. 5  is a table used to explain the operation of one preferred embodiment of the hydraulic actuator; 
       FIG. 6  is a schematic illustration of another preferred embodiment which utilizes another design of supplemental flow mechanisms; 
       FIG. 7  is a schematic illustration of another preferred embodiment which utilizes yet another design of supplemental flow mechanisms; 
       FIG. 8  depicts in more details the stroke controller of the preferred embodiment illustrated in  FIG. 7 ; 
       FIG. 9  is a schematic illustration of another preferred embodiment which utilizes yet another design of supplemental flow mechanisms; 
       FIG. 10  is a schematic illustration of another preferred embodiment which utilizes one set of rack and pinion to drive the stroke controller; 
       FIG. 11  is a schematic illustration of another preferred embodiment which utilizes two pins to drive the stroke controller; 
       FIG. 12  is a schematic illustration of another preferred embodiment which has another snubbing mechanism and uses two 3-way switch valves, instead of one 4-way switch valve; 
       FIG. 13   a  is a drawing of different alternative embodiment of the invention, including an end switch valve; 
       FIG. 13   b  is a drawing of yet a further alternative embodiment of the invention, including a differently configured end switch valve; 
       FIG. 14  is a drawing of yet a further alternative embodiment of the invention, including an end snubber valve, an extra stroke control chamber, more compact spatial arrangement of the first and second grooves, and two separate spring retainers; 
       FIG. 15  is a drawing of yet a further alternative embodiment of the invention, including a differently configured extra stroke control chamber and a first piston rod extension; 
       FIG. 16  is a drawing of yet a further alternative embodiment of the invention, including a variation in the spatial arrangement of the first and second actuation springs, which substantially overlap each other along the longitudinal axis to reduce the length of the actuator, and a variation in the spatial arrangement of the first and second grooves; 
       FIG. 17   a  is a drawing of yet a further alternative embodiment of the invention, including another variation in the design of supplemental flow mechanisms utilizing a 3-way shuttle valve, with a first flow mechanism and second-supplemental flow mechanism being open when an actuation piston overlaps with a first partial cylinder; and 
       FIG. 17   b  is a drawing of the same alternative embodiment as in  FIG. 17   a , with a second flow mechanism and first-supplemental flow mechanism being open when an actuation piston overlaps with a second partial cylinder. 
   

   DETAILED DESCRIPTION OF THE INVENTION 
   Referring now to  FIG. 1 , a preferred embodiment of the invention provides an engine valve control system using a piston, a bypass passage, and a pair of actuation spring means. The system comprises an engine valve  20 , a hydraulic actuator  30 , a high-pressure hydraulic source  70 , a low-pressure hydraulic assembly  76 , and an actuation switch valve  80 . 
   The high-pressure hydraulic source  70  includes a hydraulic pump  71 , a high-pressure regulating valve  73 , a high-pressure accumulator or reservoir  74 , a high-pressure supply line  75 , and a hydraulic tank  72 . The high-pressure hydraulic source  70  provides necessary hydraulic flow at a high-pressure P_H. The hydraulic pump  71  circulates hydraulic fluid from the hydraulic tank  72  to the rest of the system through the high-pressure supply line  75 . The high-pressure P_H is regulated through the high-pressure regulating valve  73 . The high-pressure accumulator  74  helps smooth out pressure and flow fluctuation and is optional depending on the total system capacity or elasticity, flow balance, and/or functional needs. The hydraulic pump  71  can be either of a variable- or fixed-displacement type, with the former being more energy efficient. The high-pressure regulating valve  73  may be able to vary the high-pressure value for functional needs and/or energy efficiency. 
   The low-pressure hydraulic assembly  76  includes a low-pressure accumulator or reservoir  77 , the hydraulic tank  72 , a low-pressure regulating valve  78 , and a low-pressure line  79 . The low-pressure hydraulic assembly  76  accommodates exhaust flows at a back-up or low-pressure P_L. The low-pressure line  79  takes all exhaust flows back to the hydraulic tank  72  through the low-pressure regulating valve  78 . The low-pressure regulating valve  78  is to maintain a design or minimum value of the low-pressure P_L. The low-pressure P_L is elevated above the atmosphere pressure to facilitate back-filling without cavitation and/or over-retardation. The low-pressure regulating valve  78  can be simply a spring-loaded check valve as shown in  FIG. 1  or an electrohydraulic valve if more control is desired. The low-pressure accumulator  77  helps smooth out pressure and flow fluctuation and is optional depending on the total system capacity or elasticity, flow balance, and/or functional needs. 
   The actuation switch valve  80  is a 2-position 4-way valve that supplies the hydraulic actuator  30  through a first port fluid line  192  and a second port fluid line  194 . It is 4-way because it has four external hydraulic lines: a low-pressure P_L line, a high-pressure P_H line, a first port fluid line  192  and a second port fluid line  194 . It is 2-position because it has two stable control positions symbolized by left and right blocks or positions in  FIG. 1 . Its default position is the right position secured by a return spring, and its other position is the left position forced by a solenoid. At its default or right position, the valve  80  connects the second port fluid line  194  and the first port fluid line  192  with the high pressure P_H and low pressure P_L lines, respectively. The connection order is switched when the valve  80  is at its left position. 
   The engine valve  20  includes an engine valve head  22  and an engine valve stem  24 . The engine valve  20  is mechanically connected with and driven by the hydraulic actuator  30  along a longitudinal axis  116  through the engine valve stem  24 , which is slideably disposed in the 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 air flow in/out of the associated engine cylinder. 
   The hydraulic actuator  30  comprises an actuator housing  64 , 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  68 , which is interrupted by a second-supplemental chamber  41  and a first chamber  40 , a first partial cylinder  114 , a first cavity  142 , a second cavity  144 , a third cavity  146  and a fourth cavity  148 . A stroke controller  123  resides slideably inside the first and second cavities  142  and  144 . Inside the stroke controller  123  from the first to second direction, there are a second partial cylinder  115  and a second bore  106 , which is interrupted by a second chamber  104  and a first-supplemental chamber  105 . 
   Slideably within these hollow elements of the housing  64  and the stroke controller  123  lies a shaft assembly  31  comprising, from the first to the second direction, a first piston rod  34 , a first neck  39 , a first shoulder  44 , an actuation piston  46 , a second shoulder  50 , a second neck  53 , a second piston rod  66 , and a spring seat  60 . The shaft assembly  31  further comprises a first rod passage  150  inside and across the first piston rod  34 , a second rod passage  152  inside and across the second piston rod  66 , one or more piston passages  154  inside and across the actuation piston  46 , and one or more center passages  156  inside and along the shaft assembly, interconnecting the first and second rod passages  150  and  152  and the piston passage  154 . 
   There are a first fluid space  84  defined by a cylinder first end  132  and an actuation piston first surface  92  and a second fluid space  86  defined by a cylinder second end  134  and the actuation piston second surface  98 . 
   The actuation switch valve  80  communicates with the first chamber  40  through a first port  56  and the first fluid line  192  and with the second chamber  104  through a first groove that is one or more undercuts, a second port  42 , and the second port fluid line  194 . For the purpose of easy illustration, the first and second ports  56  and  42  and their associated flow channels are in the same plane and 180-degree apart, which is not necessarily so in its physical rendition. For example, it may be physically more attractive to place them substantially on the same side of the housing  64  for easy connection with the actuation switch valve  80 . First and second grooves  108  and  109  are intended to keep, regardless the longitudinal position of the stroke controller  123  relative to the actuator housing  64 , uninterrupted fluid communication between the second chamber  104  and the second port  42  and between the first-supplemental chamber  105  and the first port  56 , respectively. The grooves  108  and  109  also help keep hydrostatic force balance on the stroke controller  123 . 
   The first cavity  142  has a substantially larger cross-section than the actuation piston  46  does, resulting in a bypass passage  48 , which provides a hydraulic short circuit between the first and second fluid spaces  84  and  86  when the actuation piston  46  does not longitudinally overlaps either of the two partial cylinders  114  and  115 . With the hydraulic short circuit, fluid may flow with substantially low resistance between the first and second fluid spaces  84  and  86 , which are thus at substantially equal pressure. The radial clearance between the first piston rod  34  and the first bore  68  and that between the second piston rod  66  and the second bore  106  are substantially small and restrictive to fluid flow. 
   Most of the design details are intended to control fluid communication between the first fluid space  84  and the first port  56  and that between the second fluid space  86  and the second port  42  through four flow mechanisms FM 1 , FM 1 S, FM 2  and FM 2 S described in details in  FIG. 2 , which, like several other figures later, does not include all parts of the actuator  30  for ease of illustration and visualization. The first flow mechanism FM 1  and the first-supplemental flow mechanism FM 1 S control fluid communication between the first fluid space  84  and the first port  56 . The first flow mechanism FM 1  runs through the first chamber  40  and the annular space between the first bore  68  and the first neck  39 , whereas the first-supplemental flow mechanism FM 1 S runs through the second groove  109 , the first-supplemental chamber  105 , the second rod passage  152 , the center passage  156 , the piston passage  154 , and the bypass passage  48 . The first flow mechanism FM 1  is open only when the actuation piston  46  longitudinally overlaps or penetrates into the first partial cylinder  114  because by design, the first piston rod  34  at least partially underlaps the first chamber  40 , thus allowing for the flow. The first-supplemental flow mechanism FM 1 S is open only when the actuation piston  46  longitudinally overlaps or penetrates into the second partial cylinder  115  because by design, the first-supplemental chamber  105  and the second rod passage  152  overlap each other, and the actuation piston  46  does not block the first partial cylinder  114 . 
   The second flow mechanism FM 2  and second-supplemental flow mechanism FM 2 S control fluid communication between the second fluid space  86  and the second port  42 . The second flow mechanism FM 2  runs through the first groove  108 , the second chamber  104  and the annular space between the second bore  106  and the second neck  53 , whereas the second-supplemental flow mechanism FM 2 S runs through the second-supplemental chamber  41 , the first rod passage  150 , the center passage  156 , the piston passage  154 , and the bypass passage  48 . The second flow mechanism FM 2  is open only when the actuation piston  46  longitudinally overlaps or penetrates into the second partial cylinder  115  because by design, the second piston rod  66  at least partially underlaps the second chamber  104 , thus allowing for the flow. The second-supplemental flow mechanism FM 2 S is open only when the actuation piston  46  longitudinally overlaps or penetrates into the first partial cylinder  114  because by design, the second-supplemental chamber  41  and the first rod passage  150  overlap each other, and the actuation piston  46  does not block the second partial cylinder  115 . 
   With the four flow mechanisms FM 1 , FM 1 S, FM 2  and FM 2 S, the first and second fluid spaces  84  and  86  are guaranteed fluid communication with the first and second ports  56  and  42 , respectively, when there is no short circuit through the bypass passage  48 . When the bypass is effective, each of the four flow mechanisms is blocked or closed, and thus each of the two fluid spaces is closed off from its respective port, preventing an open flow between two ports  42  and  56  and energy losses. These controls are valid throughout the designed stroke range of the actuator  30 , i.e. independent of the position of the stroke controller. 
   The stroke controller  123  further comprise a flange in the second direction and associated stroke controller first and second surfaces  121  and  122 . Inside the second cavity  144  and in the first direction away from the stroke controller first surface  121  is a stroke control chamber  125 . The fluid exchange in and out of the stroke control chamber  125  is primarily controlled by a stroke control pressure P_ST through a third port  43 . There also may be some internal fluid leakage or exchange between the stroke control chamber  125  and the second groove  109 . The stroke control chamber  125  is intended to help control the position of the stroke controller  123  and thus the engine valve stroke. 
   The longitudinal position of the stroke controller  123  relative the housing  64  results from the balance of the following major forces: the contact force from the actuation piston  46  to the cylinder second end  134  when they are in contact, the hydraulic static force on the cylinder second end  134  from the pressure inside the second fluid space  86 , the hydraulic static force on a bypass second edge  100 , the hydraulic static force on the stroke controller first surface  121  from the pressure inside the stroke control chamber  125 , and forces from a stroke spring  63  and a second actuation spring  58  on the stroke controller second surface  122 . The inclusion of the stroke spring  63  is optional, depending on the balance of the rest of the forces and the stroke control requirements, and it may be eliminated if the preload of the actuation spring  58  is sufficient. 
   Many of the above mentioned forces are dynamic in nature. The contact force from the actuation piston  46  to the cylinder second end  134  exists only when they are in contact. The hydraulic static force on the cylinder second end  134  changes with the pressure inside the second fluid space  86 , which alternates primarily between the system high pressures P_H and low pressure P_L and is also influenced by transient snubbing pressure. The hydraulic static force on the bypass second edge  100  varies with the pressure inside the bypass passage  48 , which stays primarily at the system high pressure P_H and experiences transient low pressure pulse during engine valve switches between the open and closed positions. The spring force from the second actuation spring  58  on the stroke controller second surface  122  varies with the extent of the compression of the second actuation spring  58 , which in turn depends on relative positions of the stroke controller  123  and the engine valve  20 . The hydraulic static force from the pressure inside the stroke control chamber  125  and the spring force from the stroke spring  63  on the stroke controller second surface  122  are independent of the engine valve movement and thus provide the stability to the position of the stroke controller  123 . The spring force from the second actuation spring  58  also has a stable component, i.e., its pre-load. The stability is further achieved by making the third port  43  fairly restrictive to fluid flow, thus damping out the high frequency oscillation caused by the engine valve switching. The third port  43  has yet to be fairly open enough to accommodate the minimum time response requirement for the stroke control. The restrictiveness of the port  43  can be replaced by another restrictive means, not shown here, between the port  43  and its fluid supply source while keeping the port  43  itself fairly open. 
   When the system power is off as shown in  FIG. 1 , the hydraulic static forces are all zero, and thus the stroke controller  123  is pushed by the springs  63  and  58  all the way against the second cavity first end  158 , when the stroke controller displacement Xst=0, and the engine valve stroke ST=STmin+Xst=STmin, with STmin being the minimum stroke and approximately equal to L 2 +L 3 , where L 2  is the depth or length of the second partial cylinder  115  as shown in  FIG. 1 , and L 3  is the overlap between the actuation piston  46  and the first partial cylinder  114  when the engine valve is fully closed as shown in  FIG. 3 . The L 3  value varies with the state of the engine valve lash, which is accommodated by having L 1 &gt;L 3  during the entire useful life of an engine. If the stroke controller  123  is pushed back all the way against the second cavity second end  160  with the stroke controller displacement Xst=STmax−STmin as shown in  FIG. 4 , not in  FIG. 1 , the engine valve has the maximum stroke ST_max i.e. the engine valve stroke ST=STmin+Xst=STmin+(STmax−Stmin)=STmax. When the power is off as in  FIG. 1 , the longitudinal distance between the stroke controller second surface  122  and the second cavity second end  160  is equal to the difference between the maximum and minimum strokes, i.e., ST_max−ST_min. 
   The continuous control of the stroke for the preferred embodiment shown in  FIG. 1  can be realized through varying the stroke control pressure P_ST by a proportional pressure control subsystem or valve (not shown here). One proportional pressure control valve can control several hydraulic actuators, for example, all intake actuators of an engine. The stroke can also be varied by actively varying the high pressure P_H while the stroke control pressure P_ST is relatively fixed, which is feasible because the required latching pressure decreases with the stroke and thus the preload of the springs. If necessary, one can regulate both P_ST and P_H, especially if P_H has to be varied for other reasons, such as energy reduction at lower strokes. 
   If the function of the continuous or proportional control of the stroke is not needed, the embodiment in  FIG. 1  can still be effectively utilized by setting P_ST at two values: a low value to have the minimum stroke and a high value for the maximum stroke or the normal full open stroke. As explained later, the minimum stroke position is necessary for the start-up of the actuator  30 . For simplicity, these two values can be simply P_H and P_L, which can be selected using a three-way valve, not shown here. 
   The first and second partial cylinders  114  and  115  have a length of L_ 1  and L_ 2 , respectively. It is intended that the actuation piston  46  will never hit the cylinder first end  132 , and its travel in the first or engine-valve-closing direction will always be stopped by the contact of the engine valve head  22  with the engine valve seat  26  when there is still a distance between the actuation piston first surface  92  and the cylinder first end  132  to accommodate the engine valve lash adjustment due to mechanical inaccuracy, wear and thermal expansion. When moving in the second direction and opening the engine valve, the actuation piston  46  stops when its second surface  98  hits the cylinder second end  134  which may not necessarily be a metal to metal contact if a proper snubbing mechanism or a squeeze film mechanism is designed. Preferably, the sum of the lengths L_ 1  and L_ 2  is substantially less than the valve stroke ST or the maximum valve stroke ST_max to minimize the loss of hydraulic energy. 
   The first and second shoulders  44  and  50  are intended to work together with the first and second bores  68  and  106  as snubbers to provide damping to the shaft assembly  31  near the end of its travel in the first and second directions, respectively. When traveling in the first direction, the actuation piston  46  pushes hydraulic fluid from the first fluid space  84  to the first chamber  40  once the actuation piston first surface  92  is distal to the bypass first edge  94 . Before the end of a stroke, the first shoulder  44  is pushed into the first bore  68 , resulting in a flow restriction because of a narrower radial clearance between the first shoulder  44  and the first bore  68  and thus a rising pressure inside the first fluid space  84  and on the actuation piston first surface  92 , which slows down the shaft assembly  31 . A similar flow restriction through the radial clearance between the second shoulder  50  and the second bore  106  helps damp the motion of the shaft assembly  31  and the engine valve  20  in the second direction. The flow restriction can be physically realized in forms other than the radial clearance. For example, notches or slots (not shown) can be cut into either the shoulders  44  and  50  or the walls of the first and second bores  68  and  106  to create desired restrictive flow openings while the clearance between the shoulders and bores are kept tight. 
   To prevent fluid starvation or cavitation, a potential negative side-effect of the above discussed restrictive or snubbing mechanisms, in the first and second fluid spaces  84  and  86  at the beginnings of the engine valve opening and closing motions, respectively, one can add, to the first and second fluid spaces  84  and  86 , additional spatial or fluid volumes that are still present, i.e., not displaced, when the actuation piston  46  is at its furthest positions in the first and second directions, respectively. These additional volumes can be, for example, substantial chamfers (not shown in  FIG. 1 ) at the opening of the first bore  68  to the first fluid space  84  and the opening of the second bore  106  to the second fluid space  86 . They can also be, but not limited to, substantial grooves or undercuts (not shown in  FIG. 1 ) on the cylinder first and second ends  132  and  134  and the actuation piston first and second surfaces  92  and  98 . These additional volumes are generally more important for the second fluid space  86  because its volume may otherwise approach to zero when the engine valve is at the open position, with the actuation piston second surface  98  in contact with the cylinder second end  134 . The added volumes may also help equalize fluid pressure within each of, not between, the two fluid spaces  84  and  86 , which is again more needed for the second fluid space  86 . The lengths of the shoulders  44  and  50  may be extended, if necessary, to maintain its effective snubbing function when the chamfers are added. 
   Concentrically wrapped around the engine valve stem  24  and the second piston rod  66 , respectively, are a first actuation spring  62  and the second actuation spring  58 . The second actuation spring  58  is supported by the stroke controller second surface  122  and the spring seat  60 , whereas the first actuation spring  62  is supported by a cylinder head surface  124  and the spring seat  60 . The spring seat  60  can also be made to function as a mechanical connection between the shaft assembly  31  and the engine valve  20  or, more specifically or locally, between the second piston rod  66  and the engine valve stem  24 . The actuation springs  62  and  58  are always under compression. They are preferably identical in major geometrical, physical and material parameters, such as stiffness, pitch and wire diameters, and free-length, such that their lengths are substantially equal and that the spring seat  60  is situated between the stroke controller second surface  122  and the cylinder head surface  124  when the springs  62  and  58  are at the neutral state, when the net spring force resulting from the two opposing spring forces is zero. 
   The shaft assembly  31  is generally under two static hydraulic forces and two spring forces. The two static hydraulic forces are the pressure forces at the actuation piston first and second surfaces  92  and  98 . The two spring forces are from the two actuation springs  62  and  58  to the spring seat  60 . Mathematically, the two spring forces can be combined as a net spring force. 
   The engine valve  20  is generally exposed to two air pressure forces on the first surface  128  and the second surface  130  of the engine valve head  22 . The hydraulic actuator  30  and the engine valve  20  also experience various friction forces, steady-state flow forces, transient flow forces, contact forces, and inertia forces. Steady-state flow forces are caused by the static pressure redistribution due to fluid flow or the Bernoulli effect. Transient flow forces are caused by the acceleration of the fluid mass. Contact forces are between the engine valve head  22  and the valve seat  26  and between the actuation piston  46  and the stroke controller  123  when these parts are in physical contact. 
   Inertia forces result from the acceleration of objects, excluding fluid here, with inertia, and they are very substantial in an engine valve assembly because of the large magnitude of the acceleration or the fast timing. 
   In  FIG. 1 , there are three seals  87 ,  88  and  89  to prevent external fluid leakages. If desired, one can also add seals to prevent internal leakages among various ports, chambers, passages, etc. If desired, one can also eliminate the seals  87 ,  88  and  89  to reduce associated frictional forces, use tolerance control to minimize the external leakages, and design proper channeling means to return unpreventable leakages back into the fluid tank. 
   Start-Up 
   When the power is off, the status of the system is substantially as that shown in  FIG. 1 . The actuation switch valve  80  is at its default or right position. The second port  42  and the first port  56  are connected to the P_H and P_L lines, respectively. The P_ST, P_H and P_L lines are all at zero gage pressure because the pump  71  is off. There is no net hydraulic force on the hydraulic actuator  30 , and there is no air force on the engine valve  20  either because the engine is not running. 
   Ignoring the frictional and gravitational forces, the stroke controller  123  is pushed by the second actuation spring  58  and the stroke spring  63  all the way in the first direction against the second cavity first end  158 . The two actuation springs  62  and  58  are compressed equally to keep force balance or to be at the neutral state. By proper longitudinally sizing or design, the actuation piston  46  and the bypass passage  48  should preferably be substantially equal in length, and the actuation piston  46  is positioned slight biased in the first direction. As a result, the actuation piston  46  slightly overlaps the first partial cylinder  114  and slightly underlaps the second partial cylinder  115 , the first rod passage  150  slightly overlaps the first-supplemental chamber  41 , the second rod passage  152  slightly underlaps the first-supplemental chamber  105 , the first piston rod  34  slightly underlaps the first chamber  40 , and the second piston rod  66  completely overlaps the second chamber  104 . As a further result, the first flow mechanism FM 1  and the second-supplemental flow mechanism FM 2 S are slightly open, while the first-supplemental flow mechanism FM 1 S and the second flow mechanism FM 2  are more restricted. The extent of the above underlapping, overlapping, opening and restriction is enhanced with the increase in lash. The engine valve  20  has an opening less than L 1 . 
   At engine start, the hydraulic pump  71  is turned on first to pressurize the hydraulic circuit. During vehicle operation, the hydraulic pump  71  is preferably driven directly by the engine. One may have to use a supplemental electrical means (not shown here) to start the hydraulic pump  71 , or to add an electrically-driven supplemental pump (also not shown). 
   At this point, the stroke control pressure P_ST is to be regulated at its minimum value so that the stroke controller  123  stays stationary and in contact with the second cavity first end  158 . The actuation switch valve  80  is still at the default or right position as shown in  FIG. 1 , and the first and second ports  56  and  42  are connected to the low and high system pressures P_L and P_H, respectively. The first and second fluid spaces  84  and  86  are therefore exposed to the low and high system pressures P_L and P_H through the first fluid mechanism FM 1  and the second-supplemental fluid mechanism FM 2 S, respectively, although the extent of their openings are limited. 
   The pressure differential between the two fluid spaces  84  and  86  will be enough to drive the actuation piston  46  in the first direction and enhance the openings in the first fluid mechanism FM 1  and the second-supplemental fluid mechanism FM 2 S, which induces a positive feedback between the shaft movement and the pressure differential until a completion of the start-up when the movement is stalled by the mechanical contact between the engine valve head  22  and the valve seat  26  as shown in  FIG. 3 . The shaft assembly  31  and the engine valve  20  will stay at that position because the differential pressure force on the piston  46  is designed to over-power the net spring return force and latch them in position. 
   The state in  FIG. 3  is the longest-lasting stable state for the engine valve  20 , which for a typical engine operation stays closed roughly ¾ of the thermodynamic cycle. For the most of the rest of the cycle, the engine valve  20  travels to the other stable state (the fully open state), stays there, and returns from it. 
   In the above description of a start-up in the first direction, the actuation piston  46  and the bypass passage  48  are substantially equal in length, and the actuation piston  46  is longitudinally positioned with a slight bias in the first direction at the beginning. It is a better starting situation. If the actuation piston  46  is longitudinally positioned with no bias at the beginning, the initial pressure and kinetic energy build-up may not be as fast, and it will still work. If the actuation piston  46  is longitudinally positioned with a slight bias in the second direction at the beginning, there will be a switch of the flow mechanisms during the start-up, from the first-supplemental flow mechanism FM 1 S to the first flow mechanism FM 1  for the first fluid space  84  and from the second flow mechanism FM 2  to the second-supplemental flow mechanism FM 2 S for the second fluid space  86 . 
   If the bypass passage  48  is materially shorter than the actuation piston  46 , there will be a fluid short circuit between two ports  42  and  56  and thus significant energy loss when the actuation piston  46  overlaps simultaneously the first and second particular cylinders  114  and  115 , thus the two rod passages  150  and  152  being connected to the second and first ports  42  and  56 , respectively and simultaneous. The start-up process may still work, although not efficiently, as long as the resulting pressure loss is not too significant. The short circuit can happen during a short-stroke operation as well as a start-up. 
   If the bypass passage  48  is materially longer than the actuation piston  46 , the start-up may experience problem if at the beginning or the neutral state, the actuation piston  46  does not overlaps any of the two partial cylinders  114  and  115 , and the first and second fluid spaces  84  and  86  are short-circuited by the bypass passage  48  and are under substantially same pressure, resulting in no driving force for the start-up. The start-up may also experience problem if at the beginning of a start-up in the first direction, the actuation piston  46  overlaps the second partial cylinder  115 , then disengages the overlap with the second partial cylinder  115  but has not possessed enough kinetic energy to jump over next short-circuiting distance. Likewise, the start-up may fail if at the beginning of a start-up in the second direction, the actuation piston  46  overlaps the first partial cylinder  114 . 
   If desired, one can also complete the start-up in the second direction or with the engine valve  20  open in the end if the actuation switch valve  80  is tuned to the left position to connect the first and second ports  56  and  42  to the P_H and P_L lines, respectively. The rest of the start-up process generally reverses what is described above. 
   Valve Opening and Closing with the Maximum Stroke 
     FIG. 5  is a table to help explain the general operation of the hydraulic actuator  30 . It can be illustrated with an example at the maximum stroke. With a maximum stroke control pressure, the stroke controller is pushed all the way in the second direction and allows for the maximum stroke as shown in  FIG. 4 . Starting from a fully closed position, with the engine valve opening Xev=0, one can start an opening stroke or travel in the second direction by switch the actuation switch  80  to the right position, connecting the first and second ports  56  and  42  with the high and low pressures P_H and P_L, respectively. The first and second fluid spaces  84  and  86  are connected to the first and second ports  56  and  42  through the first flow mechanism FM 1  (as defined in  FIG. 2 ) and the second-supplemental flow mechanism FM 2 S (as defined in  FIG. 2 ), respectively, and their respective pressures reverse polarities to the high and low pressures P_H and P_L, resulting in a net hydraulic force in the second direction, which in agreement with the net spring force releases and accelerates the shaft assembly  31  and the engine valve  20  in the second direction, opening up the engine valve  20 . The shaft assembly  31  and the engine valve  20  rapidly build up a velocity. It is a very important feature of this invention that to overcome frictional losses and engine air cylinder pressure, the net hydraulic force is in the second direction and helps the engine valve open, resulting from an additional energy contribution from the hydraulic design, which is in addition to the latch-release function. When the velocity gets to a certain level, there might be a substantial pressure drop from the P_H value in the first fluid space  84  because of snubbing by the first shoulder  44  and other restriction. The second fluid space  86  may also be at a higher pressure than P_L because of various flow restrictions. 
   Once the actuation piston  46  disengages or underlaps the first partial cylinder  114 , all four flow mechanisms FM 1 , FM 2 , FM 1 S and FM 2 S, as defined in  FIG. 2 , are blocked, and the fluid is displaced from the second fluid space  86  to the first fluid space  84  through the bypass passage  48  to accommodate the piston movement. Because of the low resistance, there is no substantial pressure difference between the two fluid spaces  84  and  86 , whereas their absolute pressure values may fall somewhere between P_H and P_L depending on the overall leakage situation. The bypass is effective when the engine valve opening Xev is between approximately L 3  and (ST–L 2 ), during which no substantial amount of hydraulic power is consumed, and the hydraulic actuator  30  is first driven and then retarded primarily by the actuation springs  62  and  58 . The potential energy stored in the springs  62  and  58  as a whole is released and continues to accelerate the hydraulic actuator  30  and the engine valve  20  until passing through the half-way point of the stroke, when the actuation springs  62  and  58  as a whole start resisting the movement in the second direction and converts the kinetic energy into the potential energy. At the half-way point of the stroke, the engine valve reaches its maximum speed. 
   Once the actuation piston  46  overlaps or engages the second partial cylinder  115  when the engine valve opening Xev is between (ST–L 2 ) and ST, the first and second fluid spaces  84  and  86  reestablish their fluid communication with the first and second ports  56  and  42  at their respective pressure values of P_H and P_L through the first-supplemental flow mechanism FM 1 S and the second flow mechanism FM 2 , respectively, resulting in a net static hydraulic force in the second direction. The bypass passage  48  is no longer effective. The net spring force continues to be in the first direction, increases with the travel, and slows down the shaft assembly  31  and engine valve  20 . 
   As the second shoulder  50  penetrates deeper into the second bore  106 , the resulting flow restriction generates a dynamic pressure rise in the second fluid space  86 , resulting in a dynamic snubbing force in the first direction to slow down the shaft assembly  31  and the engine valve  20 . The snubbing force increases with the travel and travel velocity and drops to zero when the travel stops 
   There are therefore three primary forces: the spring force in the first direction, the static hydraulic force in the second direction, and the dynamic snubbing force in the first direction. The spring force resists and slows down the engine valve opening. The static hydraulic force assists the engine valve opening, especially if there has been excessive energy loss along the way and not enough kinetic energy in the shaft assembly  31  and the engine valve  20  for them to travel all the way to a full opening. The snubbing force tends to slow down the shaft assembly  31  and the engine valve  20  if they travel too fast before the actuation piston  46  hits the cylinder second end  134  of the second partial cylinder  115 . At the full opening, i.e., the engine valve opening Xev equaling to the stroke ST, the velocity is zero, the snubbing force disappears, and the static hydraulic force is designed to be large enough to hold the engine valve  20  in place against the net spring force and other minor forces. 
   The surfaces of the cylinder first and second ends  132  and  134  and the actuation piston first and second surfaces  92  and  98  are not necessarily the flat surfaces as shown in  FIG. 1 , and they may have some taper to improve stress distribution, some shape to help squeeze-film action for impact reduction, and another shape to prevent stiction. It is also possible to design the snubber at the cylinder second end  134  in such a way that the actuation piston  46  does not hit, metal-to-metal, the cylinder second end  134  at the end of an opening stroke, at least during a dynamic operation because there is not enough to time squeeze out the trapped fluid at the location. 
   Closing the engine valve is effectively a reversal of the opening process described above. It is also described in the bottom half of the table in  FIG. 5 . It is triggered by turning the actuation switch valve  80  to its default or right position. 
   Valve Opening and Closing at Other Stroke Values 
   The opening and closing processes at other stroke values are generally the same as those at the maximum stroke. At a shorter stroke, a shorter part of the travel is covered by the bypass, and the overall spring force level and the peak travel speed decrease if the system pressure does not change. When the stroke is reduced to the minimum stroke STmin, the bypass phase disappears entirely. 
   Alternatives 
     FIG. 6  depicts an alternative embodiment of the invention. The actuator  30   e  is different from that in  FIGS. 1–4  primarily in its design of supplemental flow mechanisms FM 1 S and FM 2 S, which are no longer fabricated deep inside the shaft assembly  31   e . The first and second rod passages  150   e  and  152   e  become two circular undercuts. The stroke controller  123   e  further includes a first-supplemental chamber extension  110 , which can be a circular undercut inside the second bore  106  and distal to the first-supplemental chamber  105  in the second direction, and a third groove  111 , which is one or more undercuts distal to the second groove  109  in the second direction. The first-supplemental chamber extension  110  and the third groove  111  are in fluid communication through one or more holes in radial direction. The housing  64   e  further includes a second-supplemental chamber extension  112 , a short distance away in the second direction from the second-supplemental chamber  41 , and a fluid communication channel E-E-E, which is in fluid communication directly with the second-supplemental chamber extension  112  and the bypass passage  48  and with the first-supplemental chamber extension  110  through the third groove  111 . The third groove  111  has a longitudinal expansion enough to keep non-interruptive fluid communication between the E-E-E channel and the first-supplemental chamber extension  110 , independent of the axial position of the stroke controller  123   e.    
   With the above changes, the first and second-supplemental flow mechanisms FM 1 S and FM 2 S in  FIG. 6  are different from those in  FIG. 2 , whereas the first and second flow mechanisms FM 1  and FM 2  remain essentially the same. As shown in  FIG. 6 , the first-supplemental flow mechanism FM 1 S runs between the first port  56  and the first fluid space  84 , through the second groove  109 , the first-supplemental chamber  105 , the second rod passage  152   e , the first-supplemental chamber extension  110 , the E-E-E passage, and the bypass passage  48 . The first-supplemental flow mechanism FM 1 S is open only when the actuation piston  46  longitudinally overlaps or penetrates into the second partial cylinder  115 . 
   The second-supplemental flow mechanism FM 2 S runs between the second port  42  and the second fluid space  86 , through the second-supplemental chamber  41 , the first rod passage  150   e , the second-supplemental chamber extension  112 , the E-E-E passage, and the bypass passage  48 . The second-supplemental flow mechanism FM 2 S is open only when the actuation piston  46  longitudinally overlaps or penetrates into the first partial cylinder  114 . 
   The addition of the first and second-supplemental chamber extension  110  and  112  and the third groove  111  is to keep balance radial-direction hydrostatic forces on the shaft assembly  31   e , which may also necessitate lengthening the stroke controller  123   e  and the housing  64   e.    
     FIG. 7  depicts an alternative embodiment of the invention, in which the third groove  111  and its associated features are placed in parallel with or in between the first and second grooves  108   f  and  109   f  to save longitudinal space. Its stroke controller  123   f  is illustrated in more details in  FIG. 8 . The first, second and third grooves  108   f ,  109   f  and  111   f  are, like the earlier versions, axisymmetric for side force balance and, unlike the earlier versions, do not have enough room to have complete coverage over the entire circumference. Its flow mechanisms FM 1 , FM 2 , FM 1 S and FM 2 S are generally the same as those in the embodiment shown in  FIG. 6 , except for the first-supplemental flow mechanism FM 1 S in its spatial arrangement. The scheme used in  FIGS. 7 and 8  to arrange the grooves in parallel around the circumference can also be applied to the grooves  108  and  109  in the embodiment in  FIG. 1  to save the longitudinal space if necessary as shown in  FIG. 14 . 
   Refer now to  FIG. 9 , there is a drawing of another alternative embodiment of the invention. This alternative embodiment utilizes another design of the first and second-supplemental flow mechanisms FM 1 S and FM 2 S, which are connected to the bypass passage  48  respectively by first-supplemental and second-supplemental channels  136  and  138 . Compared with the design in  FIGS. 7 and 8 , it greatly simplifies the design, especially for the first-supplemental flow mechanism FM 1 S, and reduces internal leakage. It however requires a certain minimum amount of room in the stroke controller  123   h  and the bypass passage  48  to have an adequate cross-section size for the first-supplemental channel  136 . To make room for the first-supplemental channel  136 , the first and second grooves  108   h  and  109   h  are relocated from the stroke controller  123   h  to the housing  64   h , at substantially the same longitudinal positions though, where they are still able to keep fluid communication between the second chamber  104   h  and the second port  42  and that between the first-supplemental chamber  105   h  and the first port  56 , independent of the longitudinal location of the stroke controller  123   h . This optional relocation of a groove can be extended to other embodiments and is also applicable to the third groove  111 . 
   Refer now to  FIG. 10 , there is a drawing of another alternative embodiment of the invention. The actuator  30   u  is different from that in  FIGS. 1–4  primarily in the design of the stroke control mechanism, which is now realized by a set of rack  126  and pinion  127 . The rack  126  is solidly attached the stroke controller  123   u , which no longer has a need to form, with the housing  64   u , a stroke control chamber. For better force balance, one may choose add another set of rack  126  and pinion  127  opposite to or 180 degrees away from the one shown in  FIG. 10 . The rack  126  is substantially parallel with the axis of the stroke controller  123   u  or the actuator  30   u , and its linear displacement becomes that of the stroke controller  123   u  in either of the first and second directions. On an engine, one pinion  127  or one shaft fitted with multiple pinions, not shown here, may be designed to control a multitude of the actuator racks  126 , for example, either all intake or exhaust valve actuators on a cylinder bank. It is also possible to control the position of the stroke controller  123   u  using other mechanical means, e.g. a sliding wedge or a cam, from either the first or second direction end of the actuator  30   u.    
   Refer now to  FIG. 11 , there is a drawing of another alternative embodiment of the invention. In this embodiment, the stroke controller  123   v  is controlled via one or more pins  140 , which is further driven by a mechanical means (not shown in  FIG. 11 ), e.g. a cam or a sliding wedge. The pins  140  can either be rigidly connected to or make a simple mechanical contact with the stroke controller  123   v . If it is a simple mechanical contact, the sum of the rest of the axial forces on the stroke controller  123   v  has to be in the first direction, which can be helped by the optional stroke spring  63  if not enough preload from the actuation spring  58 . If additional force is needed in the second direction because of, for example, too much preload from the actuation spring  58 , the chamber  125   v  can be pressurized like the stroke control chamber  125  in  FIG. 1 , with additional sealing consideration between the pins  140  and the holes  141 . Otherwise, the chamber  125   v  is not pressurized by the strategic location of a seal  89   v  or generous radial clearances between the stroke controller  123   v  and the second cavity  144  and between the pins  140  and the holes  141  or a combination of both. 
   The pins  140  slideably run through pin holes  141  fabricated in the housing  64   v . The pin holes  141  are not to interfere with the first and second ports  56  and  42  and associated flow channels as shown in  FIG. 1  and are not necessarily placed in the same physical plane(s) as those ports  56  and  42  and channels. That is why the second ports  56  and  42  and associated flow channels are not illustrated in  FIG. 11 , which does not exclude their existence that is implicit for proper functions of the actuator  30   v.    
   If space allows and as another option, the pins  140  can be arranged, not shown in the figures, to push or be mechanically connected to the bypass second edge  100 , instead of the stroke controller first surface  121   v , resulting in shorter pins and holes  140  and  141 . 
   For all stroke control mechanisms disclosed above and implied otherwise, the speed of control should be appropriately regulated so that the stroke variation within a single valve switch operation is not large enough to disrupt the pendulum operation of the actuators. Coupled with frictional losses and the need to overcome engine cylinder air pressure, a large stroke increase of a distance of L 2  or more in the valve opening stroke, for example, may prevent the actuation piston  46  reaches the second partial cylinder  115  as shown in  FIG. 1 , resulting in a latching failure, because the potential energy stored in the springs at the initial time of a shorter stroke is not enough, after an intermediate step as the kinetic energy, to compress the spring to a longer distance at the later time, possible even with hydraulic energy addition in the first partial cylinder  114 . On the other hand, a large stroke reduction during a stroke may present extra energy for the snubbing mechanism to handle at the end of the stroke, causing unnecessary heavy metal impact, additional stress and unusual noises. 
   Refer now to  FIG. 12 , there is a drawing of another alternative embodiment of the invention. This embodiment is different from that in  FIG. 1  primarily in its structure in the first direction end. Instead of letting it exposed in the air, the first piston rod first end  35  is now immersed in the fluid in the enclosed first bore  68   w , which is supplied through a fourth port  45  and a first end groove  67  by a fluid supply at a pressure of P_END. The first end groove is so located longitudinally that when the engine valve  20  is near the end of its closing travel, some fluid is trapped at the end of the first bore  68   w  and can escape only through one or more notches  69  on the wall of the first bore  68   w , resulting in a snubbing action to help the engine valve  20  achieve its soft landing or impact on the valve seat  26 . This snubbing mechanism can either complement or replace the snubbing function achieved by the first shoulder  44  in the engine valve closing moment, when the speed reduction is more critical than the engine valve opening moment. The details of the snubbing mechanism, i.e., the notches  69  and the first end groove  67 , are for illustration purpose only. The snubbing function can also be achieved by other known means, e.g. replacing the notches  69  with a particular radial clearance pattern between the first piston rod  34  and the first bore  68   w  near the first direction end. 
   With the capped first bore  68   w , the first piston rod first end  35  also pumps the fluid during the rest of the opening and closing strokes and experiences a hydraulic pressure force in the second direction, the magnitude of which depends on the P_END value. This hydraulic pressure force helps the engine valve  20  overcome the cylinder air pressure during the opening stroke and resists the engine valve  20  during the closing, which is not too bad considering more favorable air pressure on the engine valve  20  during the closing. With the proper selection of the P_END value, this pumping action of the fluid is added advantage in balancing overall force and energy needs during opening and closing strokes. Ideally, the P_END value should be equal to the P_L value to save a pressure control device. Also with the capped first bore  68   w , a potential external leakage site is eliminated. 
   Refer now to  FIG. 13 , there is a drawing of another alternative embodiment of the invention. This embodiment includes an end switch valve  82   a  or  82   b , which can be arranged in two different ways as shown in  FIGS. 13   a  and  13   b , respectively. The rest of the actuator is identical to those in  FIG. 12  and is therefore omitted in the illustration. In  FIG. 13   a , the end switch valve  82   a  is used to connect the fourth port  45  either to the fluid supply P_END when the valve  82   a  is its left position or to the fluid line  192  when the valve  82   a  is at its right position. The fluid supply P_END is very similar to those described in  FIG. 12  and is for normal valve operations like opening and closing during normal combustion cycles. When the fourth port  45  is connected to the fluid line  192 , which normally carries the fluid alternating between pressure values of P_H and P_L, the first piston rod first end experiences a high hydraulic force during the entire period of a valve opening stroke and a very small hydraulic force during the closing period. This adds a big boost to the valve opening effort, which can be fruitfully utilized for compression braking used in large trucks and high-cylinder-air-pressure valve operations in air hybrid vehicle. In  FIG. 13   a , the end switch valve  82   a  is switched only for the mode change from a normal operation to, say, a compression braking operation and vice versa. The actuation switch valve or valves, which supply the fluid line  192  and are not shown in  FIG. 13   a , do the fast switching for each engine valve stroke. 
   In  FIG. 13   b , the end switch valve  82   b  is used to connect the fourth port  45  either to the fluid at pressure P_E 1  or to the fluid pressure P_E 2 . The pressures P_E 1  and P_E 2  are a lower and a higher pressure, respectively. Ideally, P_E 1  and P_E 2  are equal to P_L and P_H, respectively. During normal valve opening and closing operations, the end switch valve  82   b  stays at its left position, and the actuator  30   w  works like that in  FIG. 12 . During compression braking or other high air cylinder pressure operations, the end switch valve  82   b  is switched at the same frequency as that of the actuation switch valve, not shown here, to keep the boost force on the first piston rod first end in sync with that on the actuation piston, not shown here. In this case, the extent of the boost can be regulated by varying the time period when the end switch valve  82   b  is in its right position. 
   Referring now to  FIG. 14 , there is a drawing of another alternative embodiment of the invention. This embodiment includes an end flow control mechanism, such as an end snubber valve  208  or end flow regulator  212 , to control fluid communication between the end of the first bore  68   w  and the fourth port  45 . The end snubber valve  208  is intended to switch on and off the snubbing action of the notches  69  by being at its right and left positions, respectively. When the end snubber valve  208  is at its right position, the fluid communication between the end of the first bore  68   w  and the fourth port  45  is closed, and the notches  69  functions as an effective snubber. When the end snubber valve  208  is at its left position, the fluid communication between the end of the first bore  68   w  and the fourth port  45  is open, and there will be no substantial pressure rise at the end of the first bore  68   w  to provide the snubbing function. This option of switching on and off the snubbing function of the notches  69  is useful if one uses the notches  69  only for extra snubbing, in addition to that performed by the first shoulder  44 , to achieve ultra-low landing velocity at engine idle or other operations. Otherwise, the substantially open flow through the left position of the end snubber valve  208  disengages this extra snubbing. 
   The end flow regulator  212  has a more continuously variable nature than the end snubber valve  208  does. With the end flow regulator  212 , one can introduce a varying degree of bypassing flow between the end of the first bore  68   w  and the fourth port  45 . The end flow regulator  212  can either work with or totally replace the notches  69  in achieving a varying degree of snubbing. It may even replace the snubbing function of the first shoulder  44 . 
   The notches  69  are only one example of the snubbing mechanism design. The same snubbing function can be achieved by various known designs. For example, one can eliminate the notches  69  on the wall of the first bore  68   w  and add either taper or notches at the end of the first piston rod  34 . 
   The end snubber valve  208  and the end flow regulator  212  can be driven by either electrical or hydraulic means, not shown in  FIG. 14 . For example, the flow control means can be simply driven through a force balance between a compression spring and a surface exposed a fluid control pressure, not shown in  FIG. 14 . This control pressure can be simply the stroke control pressure P_ST or the system high pressure P_H, either of which may be at a lower value during the engine idle operation. 
   As a design option, it is also feasible for either the end snubber valve  208  or end flow regulator  212  to control the fluid communication between the end of the first bore  68   w  and, instead of the fourth port  45 , the first end groove  67 . 
   The embodiment in  FIG. 14  further includes an extra stroke control chamber  222  and an associated fifth port  220 . The extra stroke control chamber  222  provides more means to control the position of the stroke controller  123   x . Ideally the fluid communication between the extra stroke control chamber  222  and its fluid source at a pressure of P_ST 2  should be as restrictive as that between the stroke control chamber  125  and its fluid source at a pressure of P_ST to help damp overly dynamic motion of the stroke controller  123   x  during engine valve opening and closing actions. The restriction can be implemented by having either a restrictive fifth port  220  or some other orifice or restriction means between the fifth port  220  and the fluid source at a pressure of P_ST 2 . 
   The extra stroke control chamber  222  and the stroke control chamber  125  are more effective in resisting the dynamic motion of the stroke controller  123   x  in the second and first directions, respectively, due to their respective large capacities for the pressure increase caused by fluid compression. On the other hand, there is a relatively smaller room for pressure drops caused by volume expansion because of cavitation, which should be avoided in general. Like the P_ST fluid source, the P_ST 2  fluid source may not necessarily be an independently controlled fluid source, and it may be simply an existing source such as the low pressure P_L supply. 
   The embodiment in  FIG. 14  further includes first and second spring retainers  236  and  234  and associated first and second locks  240  and  238 , which are one possible variation of the spring seat  60  illustrated in earlier embodiments. The second spring retainer  234  and second lock  238  are assembled to the piston second rod end  242  to help hold the second actuation spring  58 , and the first spring retainer  236  and first lock  240  are assembled to the engine valve stem end  244  to help hold the first actuation spring  62 . After the final assembly, the piston second rod end  242  and the engine valve stem end  244  are kept in physical contact, either directly or through one or more shims  246  used to help compensate for manufacturing inaccuracy, which can also be offset by placing the shims  246  at the interface  232  between the actuator housing  64   x  and cylinder head  248 . 
   The embodiment in  FIG. 14  further includes a bypass undercut  210  at the first direction end of the first cavity  142 . The bypass undercut  210  makes it possible to reduce the diameter of the stroke controller  123   x  and thus the cross section area of the bypass second edge  100  and the hydraulic force on the stroke controller  123   x  in the second direction while still keeping or achieving a reasonable size flow area for the bypass passage  48   x . This design alternative provides another avenue to help achieve proper force balance on the stroke controller  123   x . The stroke controller  123   x  further includes design variations for the second chamber  104   x , the first groove  108   x , the first supplemental chamber  105   x , and the second groove  109   x . The first and second grooves  108   x  and  109   x  substantially overlap each other along the longitudinal axis  116  to reduce the actuator length and stagger around the circumference to avoid interference with each other. Preferably, each of the first and second grooves  108   x  and  109   x  has two or more sub-grooves, just one of which shown in  FIG. 14 , axisymmetrically distributed around the circumference for fluid force balance. The sub-grooves of the first groove  108   x  are inter-connected for fluid communication through the second chamber  104   x , and the sub-grooves of the second groove  109   x  are inter-connected for fluid communication through the first supplemental chamber  105   x . The second chamber  104   x  and the first supplemental chamber  105   x  are preferably undercuts around the whole circumference of the second bore  106 . 
   Because of the discontinuous nature of the grooves  108   x  and  109   x  around the circumference, some mechanism, such as a tube key  250 , is used to prevent the stroke controller  123   x  from drifting around the circumference and to keep proper alignment and fluid communication between the first groove  108   x  and the second port  42  and between the second groove  109   x  and the first port  56 . During the assembly, the tube key  250  can be pushed, through the second port  42  and with a press-fit with the housing  64   x , in a position as shown in  FIG. 14 , with part of it extending radially into one of the sub-grooves of the first groove  108   x . This radial extension helps limits the rotation by the stroke controller  123   x.    
   Refer now to  FIG. 15 , there is a drawing of another alternative embodiment of the invention. This embodiment further includes a first piston rod extension  214  and one or more connection orifices  252 . The first piston rod extension  214  is optional and is intended to reduce, when necessary or desirable, the surface area of the first piston rod first end  35   x  and thus the displaced fluid volume during the engine valve switch actions. 
   The connection orifices  252  are intended to provide fluid communication to the extra stroke control chamber  222 , in place of the fifth port  220 , thus eliminating the P_ST 2  fluid source when two independent stroke control fluid sources are not necessary. The connection orifices  252  are small enough to provide, working with the extra stroke control chamber  222 , damping to the stroke controller  123   x . At the same time, there still is a fluid force, for the stroke control function, from the two control chambers  125  and  222  because of their cross-section area differential although they are under the same static pressure of P_ST. 
   Refer now to  FIG. 16 , there is a drawing of another alternative embodiment of the invention. This embodiment includes a variation in the spatial arrangement of the first and second actuation springs  62   y  and  58   y , which substantially overlap each other along the longitudinal axis  116  to reduce the length of the actuator  30   y . This arrangement is accommodated by a bell-shaped second spring retainer  234   y  extending well over a smaller first spring retainer  236   y . The two actuation springs  62   y  and  58   y  are no longer identical in their physical shape, with the second actuation spring  58   y  having a larger diameter than the first actuation spring  62   y  as shown in  FIG. 16 . This physical differentiation among the springs and retainers can be easily reversed, if one prefers, to have the second actuation spring  58   y  nested inside the first actuation spring  62   y , not shown in  FIG. 16 . 
   This embodiment further includes a variation in the spatial arrangement of the first and second grooves  108   y  and  109   y , which are relocated from the stroke controller  123   y  to the housing  64   y  while still maintaining their functions to keep, regardless the longitudinal position of the stroke controller  123   y  relative to the actuator housing  64   y , uninterrupted fluid communication between the second chamber  104   y  and the second port  42  and between the first-supplemental chamber  105   y  and the first port  56 , respectively. The grooves  108   y  and  109   y  also help keep hydrostatic force balance on the stroke controller  123   y . This variation can also be applied to other embodiments. 
   While it is generally preferable to have identical actuation springs to have a symmetric pendulum, there may be other requirements and/or conditions that make it more desirable to have an asymmetric pendulum. The embodiment shown in  FIG. 16  further illustrates, for example, the option of having the engine valve  20  fully closed at the power-off state. It may be also desirable to have the forces of the actuation springs  62   y  and  58   y  biasing the engine valve  20  to the second direction to counter the cylinder air pressure force, which has a more dominant push in the first direction. This bias may also help reduce the engine valve landing speed. 
   Mathematically, the respective spring forces F 1  and F 2  from the first and second actuation spring  62   y  and  58   y  are
 
 F 2=[ F 2 o+K 2*( ST max− ST )/2]− K 2*( Xev−ST/ 2) and
 
 F 1=−[ F 1 o+K 1*( ST max− ST )/2]− K 1*( Xev−ST/ 2),
 
where a force is positive when it tends to drive the engine valve  20  in the opening or second direction. The forces F 1   o  and F 2   o  are the respective spring preloads of the first and second actuation spring  62   y  and  58   y  when the stroke ST is equal to the maximum stroke STmax and when the engine valve displacement Xev is equal to half of the stroke ST/2. K 1  and K 2  are the respective spring rates. Here the springs  62   y  and  58   y  are considered to be substantially linear and thus have constant spring rates. But a similar methodology can be applied the applications when non-linear springs are more desirables. Also, they can be applied to other embodiments not in  FIG. 16 . The total actuation spring force F is equal to the sum of F 1  and F 2 , and thus
 
 F =[( F 2 o−F 1 o )+( K 2− K 1)*( ST max− ST )/2]−( K 2+ K 1)*( Xev−ST/ 2)
 
or
 
 F=Fo−K *( Xev−ST/ 2),
 
with Fo are K being the total pre-load and spring rate, and
 
 Fo =( F 2 o−F 1 o )+( K 2− K 1)*( ST max− ST )/2 and
 
 K=K 2 +K 1.
 
The value of the total spring rate K is primarily determined according to the required natural frequency of the pendulum system, which is in turn based on the desired engine valve switch time.
 
   If, for example, it is desirable to have the engine valve  20  fully closed with a contact force of Fmino from the valve seat  26  when the power is off and when the stroke ST is at the minimum stroke STmin while adding no bias to the engine valve  20  at the maximum stroke STfnax, then one has
 
F2o=F1o,
 
 K 1=( K+ 2* F min o/ST max)/[2*(1− ST min/ ST max)], and
 
 K 2= K−K 1,
 
where if K=100,000 N/m, STmin=0.002 m, STmax=0.008 m, and Fmino=20 N, then K 1 =70,000 N/m and K 2 =30,000 N/m, i.e., with the first actuation spring rate K 1  being substantially higher than the second actuation spring rate K 2 . Only relative values of the spring preloads F 1   o  and F 2   o  are given, and their absolute values are determined with consideration of other factors, including the spring strength and length, the spring dynamics, and the need to keep continuous contact between the piston second rod end  242  and the engine valve stem end  244 , which is also true for the following example.
 
   If, in another example, it is desirable to bias the engine valve  20  to positions of Xe_mino and Xe_maxo at the minimum and maximum strokes STmin and STmax, respectively, then one has
 
( F 2 o−F 1 o )= K *( Xev _max o−ST max/2),
 
 K 1= K *( Xev _max o−Xev _min o )/( ST max− ST min)), and
 
 K 2= K−K 1.
 
If the engine valve  20  is just about to close at the minimum stroke STmin when the power is off, then let Xe_mino=0. One can let Xe_mino&gt;STmin/2 and Xe_maxo&gt;STmax/2 if the bias is intended to counter the cylinder air pressure force. For example, with STmin=0.002 m, STmax=0.008 m, K=100,000 N/m, STmin/2=0.001 m, and STmax/2=0.004 m, let Xe_mino=0.0015 m and Xe_maxo=0.0045, then K 1 =K 2 =50,000 N/m and (F 2   o −F 1   o )=50 N, i.e., with the second actuation spring preload F 2   o  being substantially higher than the first actuation spring preload F 1   o .
 
   Similarly, one can derive that with STmin=0.002 m, STmax=0.008 m, K=100,000 N/m, STmin/2=0.001 m, and STmax/2=0.004 m, then the actuation springs have to have K 1 =80,000 N/m, K 2 =20,000 N/m and (F 2   o −F 1   o )=50 N to achieve, with power-off, a force bias of 50 N in the second direction at the maximum stroke and a closed engine valve with a contact force of 30 N at the minimum stroke. 
   In all the above discussions, the first and second actuation springs  62  (or  62   y ) and  58  (or  58   y ) are each identified or illustrated, for convenience, as a single spring. When needed for strength; durability or packaging, each or anyone of the first and second actuation springs  62  or  62   y  and  58  or  58   y  may include a combination of two or more springs, nested concentrically for example. 
   Referring now to  FIGS. 17   a  and  17   b , there are drawings of another alternative embodiment of the invention. These drawings, like  FIGS. 2   a  and  2   b , do not include all parts of the actuator for ease of illustration and visualization. This embodiment includes another variation in the design of supplemental flow mechanisms, utilizing a 3-way shuttle valve  260 , which controls fluid communication from the first and second ports  56  and  42  to, through the bypass passage  48   x , the first and second fluid spaces  84  and  86 . The shuttle valve  260  includes a shuttle valve spool  261  and shuttle valve first and second bores  274  and  276 . The shuttle valve spool  261  comprises three lands, the middle one  262  of which being able to engage or overlap, along its axis, the shuttle valve first and second bores  274  and  276  to block fluid communication from the first and second ports  56  and  42  as shown in  FIGS. 17   a  and  17   b , respectively, to the bypass passage  48   x . The bypass passage  48   x  is in further fluid communication with the first and second fluid spaces  84  and  86  respectively when the actuation piston  46  is not engaged in the first and second partial cylinders  114  and  115  as shown in  FIGS. 17   b  and  17   a.    
   The longitudinal position of the shuttle valve spool  261  is controlled by pressure forces from shuttle valve first and second chambers  264  and  266  at the longitudinal ends of the shuttle valve spool  261 . The shuttle valve first chamber  264  is in fluid communication with the first port  56  through a shuttle valve first orifice  268 , and its steady state pressure is thus substantially equal to that in the first port  56 . During dynamic transitions though, there is a delay between two pressure values because of the restrictive nature of the shuttle valve first orifice  268 . There are similar geometric and physical relationships among the shuttle valve second chamber  266 , the second port  42 , and a shuttle valve second orifice  270 . 
     FIGS. 17   a  and  17   b  illustrate, respectively, two steady state conditions with the first port  56  at low and high pressures P_L and P_H, the second port  42  at high and low pressures P_H and P_L, the actuation piston  46  fully engaged in the first and second partial cylinders  114  and  115 , the shuttle valve spool  261  fully biased in the first and second directions, and the shuttle valve middle land  262  fully blocking the shuttle valve first and second bores  274  and  276 , resulting in fluid communication between the first port  56  and the first fluid space  84  through the first flow mechanism FM 1  and the first-supplemental flow mechanism FM 1 S and fluid communication between the second port  42  and the second fluid space  86  through the second-supplemental flow mechanism FM 2 S and the second flow mechanism FM 2 . The first-supplemental flow mechanism FM 1 S is open via the unblocked shuttle valve first bore  274  and the bypass passage  48   x  as shown in  FIG. 17   b , whereas the second-supplemental flow mechanism FM 2 S is open via the unblocked shuttle valve second bore  276  and the bypass passage  48   x  as shown in  FIG. 17   a.    
   During the transition from the state in  FIG. 17   a  to the state in  FIG. 17   b , the shaft assembly  31  travels in the second direction in the same or similar fashion as explained earlier as long as the first-supplemental flow mechanism FM 1 S is closed and open respectively, and the second-supplemental flow mechanism FM 2 S is open and closed respectively when the actuation piston  46  is engaged in the first and second partial cylinder  114  and  115 . Once the actuation switch valve  80  is switched from the right position to the left position, the first port  56  and thus, at lease initially, the shuttle valve first chamber  264  experience a rapid rise in its pressure from the low pressure P_L to the high pressure P_H, whereas the second port  42  and thus, at least initially, the shuttle valve second chamber  266  experience a rapid drop in its pressure from the high pressure P_H to the low pressure P_L, resulting in an directional reversal of the net pressure force on the shuttle valve spool  261  from the first direction to the second direction and thus a movement of the spool in the second direction. Because of the restrictive nature of the shuttle valve orifices  268  and  270 , the movement induces delay in rates at which the pressure values rise and drop in the shuttle valve first and second chambers  264  and  266  respectively, which can be utilized to achieve a desired time sequence or spool displacement time history so that the shuttle valve middle land  262  remains substantially underlapping the shuttle valve second bore  276  before the actuation piston  46  disengages the first partial cylinder  114  and starts substantially underlapping the shuttle valve first bore  274  before the actuation piston  46  engages the second partial cylinder  115 . The location of the shuttle valve spool  261  is not significant when the actuation piston  46  is engaged in neither of the partial cylinders  114  and  115  or in the bypass mode, which provides some design flexibility for the timing of the shuttle valve  260  when a substantial part of the actuator travel is in the bypass mode. To minimize energy loss, it is not preferable for the middle land  262  to simultaneously underlap both shuttle valve bores  274  and  276 . The timing design of the shuttle valve  260  depends more on the dynamic transition at the minimum engine valve stroke, when the movements of the shuttle valve spool  261  and the shaft assembly  31  should be substantially synchronized because the bypass time period is short or does not exist. 
   Dynamics is in a reverse order for the transition from the state in  FIG. 17   b  to the state in  FIG. 17   a . The design details in  FIGS. 17   a  and  17   b  are intended to be as an example only. They do not exclude other variations. The shuttle valve  260  may lie, for example, not in parallel with the shaft assembly  31 , and its moving part may be simply a ball, instead of a spool. The moving part may be biased by at least one spring to a default or power-off position when desired. The switch of the shuttle valve may be controlled by one or more solenoids, instead of fluid forces, to achieve better control or more functions. 
   Relative to the embodiments in  FIGS. 12 and 13 , the embodiment in  FIGS. 17   a  and  17   b  no long needs the first-supplemental and second-supplemental chambers  105  and  41  (see  FIG. 12 ), the function of the first end groove  67  (see  FIG. 12 ) is combined into the elongated first chamber  40   z , and the function of the fourth port  45  (see  FIGS. 12 ,  13   a  and  13   b ) is performed by the first port  56 . With the elimination of the first-supplemental and second-supplemental chambers  105  and  41  and the fourth port  45  (see  FIG. 12 ), this embodiment ( FIG. 17 ) is much more compact longitudinally. 
   With the first piston rod first end  35   x  exposed to the pressure at the first port  56 , which is under the high pressure P_H during the opening stroke, this arrangement in  FIG. 17 , like that in  FIG. 13   a  with the valve  82   a  in the right position, is especially suited for the actuation of an engine exhaust valve to overcome high engine cylinder pressure. 
   The actuation switch valve  80  in  FIGS. 1 ,  3 ,  4 ,  14  &amp;  15  is used for the illustration purpose only and should not be considered to be the only valve type that can be used. For example, it may be replaced by two 2-position 3-way valves  80   a  and  80   b , each of which being able to control one of the two fluid lines  192  and  194  for its connection with the high pressure P_H and low pressure P_L lines as shown in  FIGS. 12 &amp; 16 . In general, a 3-way valve is easier to manufacture than a 4-way valve. 
   One can purposely introduce a time delay between the actions of the two actuation switch valves  80   a  and  80   b  for certain functions. During the engine valve opening operation, for example, one can reduce the hydraulic energy input at the beginning of the stroke by delaying the switch of the valve  80   a  and thus keeping the first fluid space  84  at low pressure P_L a little bit longer, which may be desirable if the engine air cylinder pressure is expected to be low. Also, the switch valve  80  may be controlled by two, instead of one, solenoids, with or without return spring(s). 
   Although in many illustrations, there is one actuation switch valve for each hydraulic actuator or engine valve, this need not be the case. As many modern engines have two intake and/or two exhaust valves per engine cylinder, one actuation switch valve may simultaneously control two intake or exhaust valves on the same engine cylinder if the control strategy does not call for asymmetric opening. 
   Also in many illustrations and descriptions, the fluid medium is defaulted to be hydraulic or of 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 hydraulic actuator  30  is defaulted to be in engine valve control, and it is not limited so. The hydraulic actuator  30  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.