Patent Abstract:
Actuators, and corresponding methods and systems for controlling such actuators, 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 term “fluid” includes both liquids and gases, and the actuator may be coupled to a stem to form a variable valve actuator in an internal combustion engine, for example.

Full Description:
REFERENCE TO RELATED APPLICATION  
       [0001]     This application is a continuation-in-part of U.S. patent application Ser. No. 11/154,039, filed Jun. 16, 2005, the entire content of which is incorporated herein by reference. 
     
    
     FIELD OF THE INVENTION  
       [0002]     This invention relates generally to actuators and corresponding methods and systems for controlling such actuators, and in particular, to actuators providing independent lift and timing control with minimum energy consumption.  
       BACKGROUND OF THE INVENTION  
       [0003]     Various systems can be used to actively control the timing and lift 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.  
         [0004]     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 electro-mechanical 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.  
         [0005]     Problems with an electromechanical camless system include difficulty associated with soft-landing, high electrical power demand, inability or difficulty to control lift, 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.  
         [0006]     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.  
         [0007]     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.  
         [0008]     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.  
         [0009]     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.  
         [0010]     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.  
         [0011]     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.  
         [0012]     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.  
         [0013]     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.  
         [0014]     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.  
       SUMMARY OF THE INVENTION  
       [0015]     Briefly stated, in one aspect of the invention, one preferred embodiment of an electrohydraulic actuator comprises an actuator housing, a actuation cylinder in the actuator housing, a longitudinal axis defined by the actuation cylinder with a first and second directions, an actuation piston disposed in the actuation cylinder and moveable along the longitudinal axis in the first and second directions, and first and second ports in the actuator housing. The actuation cylinder comprises first and second ends. The actuation piston comprises first and second surfaces. One preferred embodiment further comprises a first piston rod connected to the first surface of the actuation piston and disposed slideably inside a first bearing distal to the first end of the actuation cylinder, and a second piston rod connected to the second surface of the actuation piston and disposed slideably inside a second bearing distal to the second end of the actuation cylinder, a first fluid space defined by the first end of the actuation cylinder and the first surface of the actuation piston, a second fluid space defined by the second end of the actuation cylinder and the second surface of the actuation piston, a bypass means that hydraulically short-circuits the first and second fluid spaces when the actuation piston is not proximate to either of the first or second end of the actuation cylinder, a first flow mechanism between the first fluid space and the first port, a second flow mechanism between the second fluid space and the second port, first and second actuation springs biasing the actuation piston in the first and second directions, an engine valve operably connected to the second piston rod, and one or more snubbing means.  
         [0016]     The actuation piston can be latched to the first end of the actuation cylinder, 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 bypass means 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 second end of the actuation cylinder, 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.  
         [0017]     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.  
         [0018]     In one preferred embodiment, the bypass means comprises one or more passages embedded in the housing and with openings to the fluid spaces. In an alternative embodiment, the bypass means is simply an undercut around the cylinder wall.  
         [0019]     According to the invention, the engine valve is initialized to the closed position by supply high pressure fluid to a chamber under a start piston fixed on the first piston rod. Alternatively, the engine valve is initialized to the open position by supply high pressure fluid into a chamber directly above the first piston rod. In yet another alternative embodiment, a start shaft assembly is used to selectively close and disable the bypass means so that the actuation piston and cylinder system can be directly used for its own startup. Also, by blocking the bypass means with this start shaft assembly, the actuator can be operated selectively with a much smaller lift. In another alternative embodiment, pneumatic actuation springs are used, and they may be configured to complete the initialization of the actuator either in the first or second direction.  
         [0020]     The present invention provides significant advantages over other actuators and valve control systems, and methods for controlling actuators and/or engine valves. For example, 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 ability of an alternative preferred embodiment to provide a shorter valve lift is very beneficial to achieve efficient low load operation in certain engine control strategies. 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. Certain embodiments of the present invention are able to exert additional fluid pressure force in the second direction during the bypass mode, which may be necessary in some engine exhaust valve applications.  
         [0021]     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  
       [0022]      FIG. 1  is a schematic illustration of one preferred embodiment of the hydraulic actuator and hydraulic supply system;  
         [0023]      FIG. 2  is a schematic illustration of one preferred embodiment of the hydraulic actuator, which is being initialized. For simplicity, this and rest of the illustrations do not include the hydraulic supply system;  
         [0024]      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;  
         [0025]      FIG. 4  is a schematic illustration of one preferred embodiment of the hydraulic actuator, with an opening travel just started and with the bypass not in effect;  
         [0026]      FIG. 5  is a schematic illustration of one preferred embodiment of the hydraulic actuator, with the actuator in the middle range of an opening travel and with the bypass in effect;  
         [0027]      FIG. 6  is a schematic illustration of one preferred embodiment of the hydraulic actuator, with the actuator near the end of an opening travel and with the bypass not in effect;  
         [0028]      FIG. 7  is a schematic illustration of one preferred embodiment with the engine valve fully open;  
         [0029]      FIG. 8  is a schematic illustration of another preferred embodiment which utilizes the first piston rod directly as the start mechanism. It also features tapered end surfaces of the actuation piston and cylinder;  
         [0030]      FIG. 9  is a schematic illustration of another preferred embodiment which has in the actuation cylinder one or more undercuts as the bypass;  
         [0031]      FIG. 10  is a schematic illustration of the start-up process of another preferred embodiment;  
         [0032]      FIG. 11  is a schematic illustration of the engine valve opening process of another preferred embodiment which uses a shaft assembly to block a single bypass passage;  
         [0033]      FIG. 12  is a schematic illustration of the short valve lift opening process of another preferred embodiment which uses a shaft assembly to block a single bypass passage;  
         [0034]      FIG. 13  is an alternate embodiment of the device illustrated in  FIG. 1 ;  
         [0035]      FIG. 14  is a schematic illustration of another embodiment of the invention which comprises a single piston rod and offers additional pressure force in the second direction;  
         [0036]      FIG. 15  is a schematic illustration of another embodiment of the invention which comprises one pneumatic spring and two piston rods, with the first piston rod being smaller than the second one, and offers additional pressure force in the second direction;  
         [0037]      FIG. 16  is a schematic illustration of a further alternative embodiment of the invention which comprises two piston rods, with the first piston rod primarily for additional snubbing function, and offers additional pressure force in the second direction; and  
         [0038]      FIG. 17  is a schematic illustration of a different embodiment of the invention which comprises two pneumatic springs and two piston rods, with the first piston rod being provided for additional snubbing and mechanical support, and offers additional pressure force in the second direction. 
     
    
     DETAILED DESCRIPTION OF THE INVENTION  
       [0039]     Referring now to  FIG. 1 , a preferred embodiment of the invention provides an engine valve control system using two pistons, one or more bypass passages, and a pair of 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 , an actuation switch valve  80 , and a start switch valve  82 .  
         [0040]     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.  
         [0041]     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.  
         [0042]     The actuation switch valve  80  and start switch valve  82  supply the ports of the hydraulic actuator  30  with proper flow supply lines. The start switch valve  82  shown in  FIG. 1  is a 2-position 3-way valve. It is 3-way because it has three external hydraulic lines that include two input lines, i.e., low pressure P_L and high pressure P_H, and a fluid line  190 . It is 2-position because it has two stable control positions symbolized by left and right blocks or positions in  FIG. 1 . The left position is secured by the action of a return spring when a solenoid is not energized, and it is also called the default position. The right position is secured by energizing the solenoid. At the left and right positions, the valve  82  connects the fluid line  190  with the low-pressure P_L and high-pressure P_H lines, respectively.  
         [0043]     Following the same conventions, the actuation switch valve  80  is a 2-position 4-way valve. It has four external hydraulic lines: a low-pressure P_L line, a high-pressure P_H line, a fluid line  192  and a fluid line  194 . 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 fluid lines  192  and  194  with the low pressure P_L and high pressure P_H lines, respectively. The connection order is switched when the valve  80  is at its left position.  
         [0044]     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.  
         [0045]     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 start cylinder  32 , a first bearing  68 , a first chamber  40 , a first control bore  110 , an actuation cylinder  114 , a second control bore  102 , a second chamber  104 , and a second bearing  106 . Within these hollow elements from the first to the second direction lies a shaft assembly  31  comprising a start piston  196 , a first piston rod  34 , a first shoulder  44 , an actuation piston  46 , a second shoulder  50 , a second piston rod  66 , and a spring seat  60 . The first piston rod  34  further comprises a first-piston-rod second neck  38 , a first land  90 , and a first-piston-rod first neck  39 . The second piston rod  66  further comprises a second-piston-rod first neck  53 , a second land  52 , and a second-piston-rod second neck  54 .  
         [0046]     In the actuation cylinder  114 , there is a first fluid space  84  defined by the actuation cylinder first end  132  and the actuation piston first surface  92  and a second fluid space  86  defined by the actuation cylinder second end  134  and the actuation piston second surface  98 .  
         [0047]     The shaft assembly  31  can be substantially radially supported by some or all of the following mating surfaces from the first to the second direction: the start piston  196  and the start cylinder  32 , the first piston rod  34  and the first bearing  68 , the actuation piston  46  and the actuation cylinder  114 , and the second piston rod  66  and the second bearing  106 . Each pair of the above listed mating surfaces has tight clearance, provides substantial hydraulic seal, and yet offers tolerable resistance to relative motions, including translation along and, if desired, rotation around the longitudinal axis  116 , between the shaft assembly  31  and the housing  64 . The start cylinder  32  communicates hydraulically with the start switch valve  82  through a start port  36  and the fluid line  190 . The actuation switch valve  80  communicates with the first chamber  40  through a first port  42  and the fluid line  192  and with the second chamber  104  through a second port  56  and the fluid line  194 .  
         [0048]     Through the side wall of the actuation cylinder  114 , there are one or more bypass passages  48 , which provide a hydraulic short circuit over a substantial length of the actuation cylinder  114 . The bypass passages  48  are preferably arranged in such a way that there is on the actuation piston  46  minimum net side force due to hydraulic static pressure. With the hydraulic short circuit, fluid may flow with substantially low resistance between the first and second fluid spaces  84  and  86 , and the entire actuation cylinder  114  is at substantially equal pressure. The hydraulic short circuit is not effective either when the actuation piston first surface  92  is distal, in the first direction, to the bypass first edge  94  or the actuation piston second surface  98  is distal, in the second direction, to the bypass second edge  100 . The longitudinal distance between the bypass first edge  94  and the actuation cylinder first end  132  is L_ 1 . The longitudinal distance between the bypass second edge  100  and the actuation cylinder second end  134  is L_ 2 .  
         [0049]     The first land  90 , the first control bore  110 , and the first-piston-rod first and second necks  39  and  38  work together as a flow mechanism. The first land  90  selectively blocks fluid flow between the first chamber  40  and the first fluid space  84  of the actuation cylinder  114 , which occurs when the first land  90  is longitudinally located in or overlaps the first control bore  110 , with the radial clearance between the first land  90  and the first control bore  110  being substantially small and restrictive to fluid flow. The second land  52 , the second control bore  102 , and the second-piston-rod first and second necks  53  and  54  work together as another flow mechanism. The second land  52  selectively blocks fluid flow between the second chamber  104  and the second fluid space  86  of the actuation cylinder  114 , which occurs when the second land  52  is longitudinally located in or overlaps the second control bore  102 , with the radial clearance between the second land  52  and the second control bore  102  being substantially small and restrictive to fluid flow.  
         [0050]     The longitudinal locations of the first land  90  and the second land  52  along the shaft assembly  31  are such that each of the two lands  90  and  52  blocks fluid flow when the actuation piston  46  sits or travels in-between the bypass first and second edges  94  and  100 , i.e., the bypass passages  48  being in effect. This prevents an open flow, through the bypass passages  48 , between the first chamber  40  and the second chamber  104  and saves energy. When the bypass passages  48  are not effective, the two lands  90  and  52  disengage or underlap their respective control bores  110  and  102  and allow substantial flow between the first chamber  40  and the first fluid space  84  and between the second chamber  104  and the second fluid space  86 .  
         [0051]     The lengths of the actuation piston  46  and cylinder  114  are designed such that the piston  46  can travel with a stroke of ST plus an allowance for the engine valve lash adjustment. When moving in the second direction and opening the engine valve, the actuation piston  46  stops when its second surface  98  hits the actuation cylinder second end  134 . When moving in the first direction and closing the engine valve, the engine valve head  22  hits the valve seat  26  first while there is still a distance L_lash (see  FIG. 3 ) or less between the actuation piston first surface  92  and the actuation cylinder first end  132 . The distance L_lash is allowance for the engine valve lash adjustment. Preferably, the sum of the lengths L_ 1  and L_ 2  is substantially less than the valve stroke ST to minimize the loss of hydraulic energy.  
         [0052]     The first and second shoulders  44  and  50  are intended to work together with the first and second control bores  110  and  102  as snubbers to provide damping of the shaft assembly  31  near the end of the 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 . At roughly the same time, the first shoulder  44  is pushed into the first control bore  110 , resulting in a flow restriction because of a narrower radial clearance between the first shoulder  44  and the first control bore  110  and thus a rising pressure on the actuation piston first surface  92 , which slows down the shaft assembly. A similar flow restriction through the radial clearance between the second shoulder  50  and the second control bore  102  helps dampen the motion of the shaft assembly  31  and the engine valve  20  in the second direction.  
         [0053]     Concentrically wrapped around the engine valve stem  24  and the second piston rod  66 , respectively, are a first actuation spring  62  and a second actuation spring  58 . The second actuation spring  58  is supported by the housing surface  122  and the spring seat  60 , whereas the first actuation spring  62  is supported by cylinder head surface  124  and spring seat  60 . 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 the net spring force resulting from the two opposing spring forces is substantially equal to zero at the neutral position shown in  FIG. 1 .  
         [0054]     The spring seat  60  is designed such that when it is located substantially half-way between the housing surface  122  and the cylinder head surface  124  and when the actuation piston  46  is at the longitudinal center of the actuation cylinder  114  as shown in  FIG. 1 , the two actuation springs  62  and  58  are under equal compression. As such the net spring force is zero, which is also the neutral position of the hydraulic actuator  30 , with the engine valve  20  being open at half of its stroke ST. The spring seat  60  also offers 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 .  
         [0055]     The shaft assembly  31  is generally under three static hydraulic forces and two spring forces. The three static hydraulic forces are the pressure forces at the actuation piston first and second surfaces  92  and  98  and the start piston second surface  127 . The start piston first surface  126  is preferably exposed to the air or a low pressure fluid. In case of a hydraulic leakage around the start piston  196 , a passage may be included to channel the leak flow from the top of the piston  196  to the hydraulic tank. The two spring forces are from the two actuation springs  62  and  58  to the spring seat  60 .  
         [0056]     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, 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. 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.  
       Start-Up  
       [0057]     When the power is off, the status of the system is substantially equal to that shown in  FIG. 1 . Two switch valves  80  and  82  are at their default positions. The start port  36  is connected to the P_L line, and the first port  42  and the second port  56  are connected to the P_L and P_H lines, respectively. Both the P_H and P_L lines are 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.  
         [0058]     Ignoring the gravitational force, the two springs  62  and  58  have to be compressed equally to keep force balance, resulting in a longitudinally centered position for the spring seat  60  between the housing surface  122  and the cylinder head surface  124 , a longitudinally centered position for the actuation piston  46  in the actuation cylinder  114 , and a half-open position for the engine valve  20 .  
         [0059]     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).  
         [0060]     Even with the system pressurized, however, the actuation piston  46  is stationary because its two surfaces  92  and  98  are exposed to substantially the same pressure due to the bypasse(s)  48 . Instead, the start switch valve  82  has to be turned to its start or right position as shown in  FIG. 2 , with the second surface  127  of the start piston  196  being exposed to the high pressure P_H. The start piston  196  thus pulls, in the first direction, the shaft assembly  31  and the engine valve  20 , overcoming the net spring force. Note that the actuation switch valve  80  is still in its default or right position as shown in  FIG. 2 , and it supplies the first chamber  40  and the second chamber  104  with the low pressure P_L and high pressure P_H lines, respectively.  
         [0061]     Once the actuation piston first surface  92  travels past the bypass first edge  94 , the bypass passages  48  are blocked or disabled, and flows through the first and second control bores  110  and  102  are no longer blocked by the first and second lands  90  and  52 , resulting in a driving force in the first direction on the actuation piston  46  with the high pressure P_H and low pressure P_L at its second and first surfaces  98  and  92 , respectively. This differential pressure force is set to be strong enough to hold the shaft assembly  31  and the engine valve  20  in the closed position against the spring force even after the start switch valve  82  is switched back to its default or non-start position and supplies only low pressure P_L fluid to the start cylinder  32  as shown in  FIG. 3 .  
         [0062]     At the state shown in  FIG. 3 , the start-up process is complete, start switch valve  82  will remain in the default or non-start or left position until the next engine starting, and the start cylinder  32  will remain filled with low-pressure fluid and contribute negligible force to hydraulic actuator  31 . Due to the back-and-forth movements of the start piston  196  during the normal operation, the pressure inside the start cylinder  32  deviates from the system low-pressure P_L. To prevent unnecessary losses, this deviation can be minimized by having shorter and larger flow passages in the fluid line  190  and the start switch valve  82 . The time response requirement for the start-up is generally not as stringent as that for the engine valve switching, the start switch valve  82  can be made with larger openings.  
         [0063]     The state in  FIG. 3  is a 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.  
       Valve Opening  
       [0064]     To open the engine valve  20 , the actuation switch valve  80  is turned to the left position as shown in  FIG. 4 , wherein the first and second chambers  40  and  104  are connected with the high pressure P_H and low pressure P_L, respectively. Due to the open communication through the second control bore  102 , the pressure in the second fluid space  86  quickly drops close to the low pressure P_L. Although the first control bore  110  is somewhat restricted by the first shoulder  44 , the pressure in the first fluid space  84  still can reach close to the high pressure P_H within a reasonable amount of time because of a low initial piston speed and flow rate. With these actuations, the differential hydraulic force on the actuation piston  46  changes its direction from in the first direction to in the second direction. This hydraulic force in the second direction works with the net spring force in the same direction to accelerate the shaft assembly  31  and the engine valve  20 , and also helps overcome whatever engine cylinder air force on the engine valve head  22 .  
         [0065]     When the engine valve opening is between (L_ 1 -L_lash) and (ST-L_ 2 ) during the travel in the second direction as shown in  FIG. 5 , the first and second control bores  110  and  102  are substantially blocked by the first and second lands  90  and  52 , respectively, and the displacement of the actuation piston  46  is accomplished by flows through the bypass passages  48 . Hydraulic power is no longer used, and the hydraulic actuator  31  is driven primarily by the actuation springs  62  and  58 . The potential energy stored in the springs  62  and  58  is released and continues to accelerate the hydraulic actuator  31  and the engine valve  20  until passing through the half-way point of the stroke, when the actuation springs  62  and  58  start resisting the movement in the second direction and converts the kinetic energy into the potential energy.  
         [0066]     When the engine valve opening is between (ST-L_ 2 ) and ST during a travel in the second direction as shown in  FIG. 6 , both the first and second control bores  110  and  102  are open for flows. Within this travel range, the net spring force is in the first direction, increases with the travel, and slows down the shaft assembly  31  and engine valve. When the actuation piston second surface  98  just passes the bypass second edge  100 , the first and second surfaces  92  and  98  of the actuation piston  46  are now exposed to the high pressure P_H and low pressure P_L, respectively, resulting in a net static hydraulic force in the second direction.  
         [0067]     As the second shoulder  50  penetrates deeper into the second control bore  102 , 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  
         [0068]     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 actuation cylinder  114 . At the full opening as shown in  FIG. 7 , the snubbing force disappears, and the static hydraulic force should be large enough to hold the engine valve  20  in place against the net spring force and other minor forces.  
       Valve Closing  
       [0069]     Closing the engine valve is effectively a reversal of the opening process just described. It is triggered by turning the actuation switch valve  80  to its default or right position as shown in  FIG. 3 . Upon completion, the hydraulic actuator  30  and the engine valve  20  are back to their default states as shown in  FIG. 3 .  
         [0070]      FIG. 8  depicts an alternative embodiment of the invention. The primary physical difference between this embodiment and that illustrated in  FIGS. 1 through 7  lies in the start-up mechanism. This alternative configuration does not include a start piston, but instead utilizes a combination of the first piston rod  34  and a new first bearing  68   b , which is more extended longitudinally than the first bearing  68  in  FIGS. 1-7 .  
         [0071]     In operation, the start switch valve  82  is turned to its start or right position as shown in  FIG. 8  and supplies the high pressure P_H fluid to the first bearing  68   b , resulting in a hydraulic force on the first-piston-rod end surface  136 , which pushes the shaft assembly  31   b  and the engine valve  20  to the full open position. To complete the initialization, the actuation switch valve  80  has to be turned to its left position as shown in  FIG. 8  so that the first and second chambers  40  and  104  are supplied with the high pressure P_H and low pressure P_L fluids, respectively.  
         [0072]     Once the start-up is complete, this embodiment operates like the embodiment in  FIGS. 1 through 7 . This alternative embodiment has a simpler starting mechanism, but application may be limited by the available space between the fully-opened engine valve  20  and the top of the engine piston at the top dead center to avoid physical interference or impact. This embodiment also features tapered end surfaces for the actuation piston  46   b  and actuation cylinder  114   b . When the actuation piston second surface  98   b  hits the actuation cylinder second end  134   b , the tapered surfaces may have better stress distribution and longer service life. Although in a preferable design, the actuation piston first surface  92   b  will never hit the actuation cylinder first end  132   b , still their tapered shape may help release local stress caused by high snubbing pressure. To achieve the same flow blocking function and logic, the first and second lands  90   b  and  52   b  are extended in their lengths compared with the lands in other preferred embodiments.  
         [0073]     Refer now to  FIG. 9 , there is a drawing of another alternative embodiment of the invention. The main physical difference between this embodiment and that illustrated in  FIGS. 1 through 7  lies in the design of the bypass in the actuation cylinder  114 . In this embodiment, the bypass is one or more bypass undercuts  138 . This design provides smoother or freer bypass flow around the actuation piston  46  between the first and second edges  94   b  and  100   b  and less friction on the piston  46 .  
         [0074]     Refer now to  FIG. 10 , which is a drawing of yet another alternative embodiment of the invention. Compared with the embodiment in  FIG. 8 , this embodiment is different primarily in its start mechanism  150 , which is designed to block a bypass passage  152 , preferably the only bypass passage around the actuation cylinder  114 . Also, the shaft assembly  31   d  does not include the first land  90   b  as in  FIG. 8 , resulting in an extended neck  389 . The reason for the elimination of the first land  90  will become clear when the operation of this embodiment is explained below.  
         [0075]     The start mechanism  150  includes a start shaft  154  comprising a first head  156 , a second head  160  and a stem  158  in between the two heads  156  and  160 . The start shaft  154  moves inside the bypass passage  152 , which is extended longitudinally beyond the length necessary for the bypass flow function to accommodate the whole length of the start shaft  154 . Two ends of the bypass passage  152  are hydraulically connected to start first and second ports  162  and  164 , respectively. Between the bypass passage  152  and the start first port  162 , there is a smaller passage  166 , offering a limit shoulder  140  to offer the limit in the first direction for the movement of the start shaft  154 . A return spring  168  resides inside the small passage  166  and, when the start shaft  154  is not all the way against the limit shoulder  140 , a part of the bypass passage  152  to urge the start shaft towards the second direction. The start first port  162  is always connected with the low pressure P_L line, whereas the start second port  164  is connected with either the high pressure P_H or low pressure P_L lines through the start switch valve  170 .  
         [0076]     The bypass passage  152  and the start shaft  154  have a reasonable radial clearance to ensure a smooth sliding movement for the shaft  154  and minimum hydraulic leakage. From the first to the second direction along the longitudinal axis of the bypass passage  152 , there are a first bypass groove  172 , a second bypass groove  174  and a check valve groove  176 . From the first to the second direction along the longitudinal axis of the actuation cylinder  114 , there are a first actuation cylinder groove  178  and a second actuation cylinder groove  180 . These five grooves are intended to reduce or eliminate hydraulic force imbalance on the start shaft  154  and the actuation piston  46  and to facilitate the reduction of the flow resistance. The first bypass groove  172  is in hydraulic communication with the first actuation cylinder groove  178 , whereas the second bypass groove  174  is in hydraulic communication with the second actuation cylinder groove  180 . The check valve groove  176  is in hydraulic communication, C-to-C, with the downstream side of a check valve  182 , whereas the upstream end of the check valve  182  is in hydraulic communication with the second port  56  or, not shown in  FIG. 10 , with the second chamber  104 .  
         [0077]     In start operation as shown in  FIG. 10 , the start switch valve  170  is energized and set at the left position, connecting the start second port  164  to the low pressure P_L line. The start shaft  154  is pushed by the return spring  168  in the second direction and blocks, with the first head  156 , the first bypass groove  172  and the bypass passage  152 , and the actuation piston  46  functions like a normal piston. Also, the actuation switch valve  80  is in its default or right position, connecting the first and second ports  42  and  56  to the low pressure P_L and high pressure P_H lines, respectively. The first fluid space  84  is now exposed the low pressure P_L because it is in hydraulic communication with the first port  42  though the first chamber  40  and the first control bore  110 , which is not blocked by the first land  90   b  as in  FIG. 8 .  
         [0078]     Although the second control bore  102  is blocked by the second land  52 , the second fluid space  86  is still exposed to the high pressure P_H because it is in hydraulic communication with the second port  56  through the check valve  182 , the hydraulic communication C-to-C, the check valve groove  176 , a portion of the bypass passage  152 , the second bypass groove  174 , and the second actuation cylinder groove  180 . The resulting differential pressure pushes the actuation piston  46  and thus the shaft assembly  31   d  and engine valve  20  all the way to the fully closed position, which completes the start-up process. Near the end of this travel, the second land  52  slides out the second control bore  102  to further ensure the connectivity between the second fluid space  86  and the second port  56 .  
         [0079]     In normal operation as shown in  FIG. 11 , the start switch valve  170  is de-energized and returned to its default or right position to keep the start second port  164  pressurized and to hold the start shaft  154  against the returning spring  168 , resulting in a substantially open bypass passage  152  and a blocked check valve groove  176 , which disables the check valve  182 . Thus, hydraulic actuator  31   d  in  FIG. 11  functions much like the hydraulic actuator  31   b  in  FIG. 8 , except that in  FIG. 11  there is only one blocking land, the second land  52  to block the free flow between the first and second ports  42  and  56  during the middle portion of a stroke when the bypass passage  152  is open.  
         [0080]     In an engine valve opening stroke as illustrated in  FIG. 11 , the actuation switch valve  80  is de-energized or at its left position and connects the first and second ports  42  and  56  to the high pressure P_H and low pressure P_L lines, respectively, and the actuation piston  46  has moved to the middle range of the movement in the second direction where the bypass passage  152  is open. At this point, the entire actuation cylinder  114  is exposed to high pressure P_H through the bypass passage  152  and first control bore  110 . The net hydraulic force on the actuation piston  46  is still equal to zero. Therefore, the elimination of the first land  90  or  90   b  does not fundamentally change the function of the system although it may introduce a little more flow leakage between the first and second ports  42  and  56  because it eliminates one of the two main barriers in the flow path. It is also workable to eliminate the first land  90  or  90   b  in other preferred embodiments in  FIGS. 1-9 .  
         [0081]     This latest embodiment is also able to drive the engine valve  20  with a small lift, which is a great plus for engine calibration and control strategy. As shown in  FIG. 12 , the actuation switch valve  80  is at its left position, and the hydraulic assembly  31   d  is in a travel in the second direction. However, the start switch valve  170  is at its left position, and the start shaft  154  is at its lower position, blocking the bypass passage  152 .  
         [0082]     As shown in  FIG. 12 , the actuation piston  46  has just traveled a distance of (L_ 1 -L_lash), and the second land  52  is about to enter the second control bore  102 . At this point, the second fluid space  86  is a closed or trapped volume, without hydraulic communication with anyone of the ports  42  and  56 . Any further motion in the second direction by the actuation piston  46  will cause a volume reduction and pressurization. The total piston travel is thus limited, barring any severe leakage, to not too much more than (L_ 1 -L_lash).  
         [0083]     Once the actuation switch valve  80  is turned to the right position and connects the first and second ports  42  and  56  to low pressure P_L and P_H lines, respectively, the high pressure fluid will enter the closed second fluid space  86  through the check valve  182  and the C-to-C connection. Shortly after that, the second land  52  is out of the second control bore  102 , and the high-pressure fluid can flow more freely into the second fluid space  86  and complete the return stroke, against the spring force, which intends to push the assembly to the neutral or middle position. During this short lift operation, the two springs  62  and  58  cannot contribute much, and entire operation has to be sustained by the hydraulic system, which is still feasible because of the shorter stroke.  
         [0084]     Various switch valves  80 ,  82 , and  170  are used for the illustration purpose only and should not be considered to be the only valves that can be used. For example, the actuation switch valve  80  may be replaced by two 2-position 3-way valves  80   a  and  80   b , each of them 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  FIG. 13 . In general, a 3-way valve is easier to manufacture than a 4-way valve.  
         [0085]     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 chamber  40  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, either or both of the two switch valves  80  and  82  may be controlled by two, instead of one, solenoids. If necessary, some of these switch valves may be controlled by pilot valves. This flexibility in valve selection applies to other preferred embodiments as well.  
         [0086]     Although in each of the illustrations so far, there is one start switch valve and 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. One start switch valve may control all the engine valves in an entire engine.  
         [0087]     With continuing reference to the drawings,  FIG. 14  illustrates another embodiment of the invention. A main feature of this actuator, depicted generally at  30   j , is the lack of a first piston rod. In this case, the first flow mechanism comprises a first control bore  110   j  which is always open for fluid communication between the first port  42  and the first fluid space  84  (except for the snubbing action when it is substantially restricted by the first shoulder  44 ). There will still be no open flow between the first and second ports  42  and  56 , because its second flow mechanism retains the second piston rod  66  and the associated second land  52  and is able to substantially block fluid communication between the second port  56  and the second fluid space  86 .  
         [0088]     With only one piston rod, the effective pressure exposure area is greater on the actuation piston first surface  92  than on the actuation piston second surface  98 , when considering the exposed area left open by the missing first piston rod. As a result, there is a net pressure force in the second direction during the bypass stage of a travel, and this net pressure force is especially significant during a travel in the second direction when the first port  42  and thus both the first and second fluid spaces  84  and  86  are at the system high pressure P_H.  
         [0089]     When traveling through the bypass mode in the first direction, the first port  42 , and thus both the first and second fluid spaces  84  and  86 , are at the system low pressure P_L, and the net pressure force is still in the second direction but relatively small. This embodiment may be used as an actuator for engine exhaust valves with significant engine cylinder air pressure force, against which a significant, asymmetric force is needed. In many cases such as exhaust valves of large two-stroke marine diesel engines, this additional force is as great as, if not more than, the force needed for engine valve acceleration.  
         [0090]     The above discussed asymmetrical area arrangement and net pressure force can also be utilized to start the actuator by switching the actuation switch valve, which doubles as a start switch valve, to its left block or position as shown in  FIG. 14 , applying a high system pressure P_H to the first port  42 . The resulting net fluid pressure force pushes the engine valve  20  to the fully open position and initialize the actuator  30   j.    
         [0091]     If the actuator has to be initialized to a fully closed position, a separate starting mechanism can be incorporated. For example, a mechanism such as that illustrated in  FIGS. 10-12  can be used to temporarily block the bypass passage for an effective initialization in the first direction.  
         [0092]     The embodiment of  FIG. 14  comprises an optional first snubber check valve  142 , which helps backfill and reduce potential cavitation in the first fluid space  84  at the beginning of travel in the second direction. The first snubber check valve  142  allows for flow from the first port  42  or the first control bore  110   j  (not shown in  FIG. 14 ) to the first fluid space  84 , but not in the opposite direction. Similar snubber check valves can be applied to other snubbers of this invention when desired and practical. The illustration in  FIG. 14  is more as a symbol than the actual design form of a check valve. Such valves can incorporate, for example, a ball with a preload spring or a reed. In general, these check valves should exhibit a fast dynamic response. In situations where an appropriate check valve is not available, it is preferable for the snubber to have a reasonable minimum fluid volume and a rational minimum orifice or opening area.  
         [0093]     The embodiment of  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  shown in earlier embodiments. The second spring retainer  234  and second lock  238  are assembled to the piston second rod end  242 . The assembly helps hold the second actuation spring  58 . The first spring retainer  236  and the 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 (not shown) to help compensate for manufacturing inaccuracy.  
         [0094]      FIG. 15  shows another alternative embodiment of the invention. This actuator, depicted generally at  30   k , includes a first piston rod  34   k , its diameter being substantially smaller than that of the second piston rod  66 , resulting in a net pressure force in the second direction during the bypass stage of a travel. This is functionally similar to that of the actuator  30   j  illustrated in  FIG. 14 , although most likely with a relatively smaller net or asymmetric force because of the presence, however small, of the cross section area of the first piston rod  34   k.    
         [0095]     The actuator  30   k  in  FIG. 15  can be initialized in ways akin to those of actuator  30   j  in  FIG. 14  due to the similar asymmetric fluid actuation design. The actuator  30   k  may be used in situations where an exhaust valve experiences relatively lower engine cylinder air pressure. Still, with the first piston rod  34   k  supported in radial direction, it is more feasible for the actuator  30   k  to adopt a simple undercut as its bypass passage  138 . Its first flow mechanism comprises the first control bore  110   k , which is not sufficiently restricted by the first piston rod  34   k  with a smaller diameter. The fluid communication between the first port  42  and the first fluid space  84  is always open except for the snubbing action, when it is substantially restricted by the first shoulder  44 . The second flow mechanism is identical to that of the embodiment in  FIG. 14 , and is able to close during the bypass mode.  
         [0096]     In the embodiment illustrated in  FIG. 15 , the second actuation spring  58  is a pneumatic spring, wherein a pressurized volume of gas is enclosed in a pneumatic cylinder  254  and a pneumatic piston  250  including an optional pneumatic piston seal  252 . The design of the pneumatic spring can be optionally replaced by other common variations, such as a bladder type of construction (not shown in  FIG. 15 ) for better leakage prevention. The pneumatic cylinder  254  can be fabricated inside the housing  64   k  (as shown in  FIG. 15 ) or in a separate mechanical block. For leakage compensation, spring force curve control, optional initialization, and other functions, the second actuation spring  58  is connected through a pneumatic port  264  and a pneumatic valve  268 , with one or more gas supplies, for example high pressure P_H_gas and low pressure P_L_gas supplies. The low pressure P_L_gas supply may not be needed in some applications, especially if the gas used is simply air. In certain applications, the pneumatic valve  268  may be replaced by a pneumatic pump (not shown in  FIG. 15 ), pumping directly from a low-pressure gas supply.  
         [0097]     The force curve control includes regulating and/or changing, in real time per functional needs and operational conditions, the force curve of the second actuation spring  58  relative to the fixed force curve of the first actuation spring  62  to achieve a desired asymmetric net spring force. This can be used, for example, to generate a load-dependent force biased on average in the second direction to help move against the engine cylinder air pressure. The real-time adjustment may be also needed for temperature compensation because of the temperature sensitive gas properties.  
         [0098]     The second actuation space  58  may be set at a low pressure or force so that the engine valve stays at or returns to the closed position because of a stronger force from the first actuation spring  62  when the engine is off, which may be a beneficial function by itself for many applications and will also help set the actuator for a proper initialization. At the next engine start, one can initialize the actuator  30   k  first by turning the actuation switch valve  80  to the right position or block as shown in  FIG. 15 , then pressuring the second actuation spring  58 .  
         [0099]     The actuator  30   k  may include a normally-open pneumatic valve  266  for applications where seating of engine valves is absolutely necessary, for example, to avoid hitting engine pistons, when the engine is off or when the electrical system is interrupted. When the solenoid is on, the normally-open pneumatic valve  266  stays at the right position, in a closed condition, and does not contribute to actuator operation. When the solenoid is off, valve  266  is driven by a return spring to the left position, opening the pneumatic port  264  to a low pressure supply (as shown in  FIG. 15 ), or directly to atmosphere (not shown), and secures the return of the engine valve to its seating position. The normally-open pneumatic valve  266  can be eliminated if its function can be incorporated in the pneumatic valve  268 .  
         [0100]     The actuator  30   k  may include an optional pneumatic bleed hole  256  to relieve the pressure on the back or non-functional side of the pneumatic piston  250  in case of an otherwise air-tight design as implied in  FIG. 15 . If desired, the second actuation spring  58  can also be located between the first actuation spring  62  and the actuation piston  46 . This pneumatic spring concept and its variations may be applied to other embodiments of this invention as well, including the example shown in  FIG. 17 . Most of other embodiments may also adopt another concept used in this embodiment: placing the two actuation springs, whether they are mechanical or pneumatic type, at the two longitudinal sides of the actuation piston.  
         [0101]      FIG. 16  shows yet a further alternative embodiment of the invention. The actuator, labeled  30   m , is a variation of the actuators  30   j  and  30   k  from  FIGS. 14 and 15 . Like the actuator  30   k , it possesses a fist piston rod  34   m ; however, it does not provide substantial mechanical support in a radial direction, and is intended to work with the dead-ended first bearing  68   m  and associated one or more notches  69  as an end snubber, functional when travel approaches the end of the first direction. At the remainder of the travel or positions, the first piston rod  34   m  is not close to being supported, and the first-piston-rod end surface  136   m  is exposed to the pressure at the first port  42 . As a consequence, the pressure force distribution is very much like that experienced by the actuator  30   j  in  FIG. 14 .  
         [0102]     Like actuator  30   j , actuator  30   m  is effective to drive a load, such as an exhaust engine valve, with asymmetric load needs in the first and second directions. With the added end snubber, it provides better control over valve seating velocity. When desired, an end snubber valve  208  may be used and turned on to deactivate the end snubber by opening fluid communication between the dead-ended first bearing  68   m  and the first port  42 , thus equalizing pressure. This function is useful in keeping two engine valve seating velocities for idle and wild-open-throttle operations, respectively, if other parameter control methods are not sufficient. If more precise, or continuously variable, control is desired, an end flow regulator  212  may be used to continuously regulate the extent of the fluid communication between the dead-ended first bearing  68   m  and the first port  42 . Either of the end snubber valve  208  and the end flow regulator  212  can be controlled or actuated externally or within the actuator itself by using an existing signal such as the system high pressure P_H.  
         [0103]      FIG. 17  shows yet a different alternative embodiment of the invention. In this embodiment, the first piston rod  34   n  works with the dead-ended first bearing  68   n  and associated one or more notches  69  as an end snubber, provides mechanical support in radial direction by being received in the first bearing  68   n  over the entire range of travel. The embodiment also offers, in the bypass mode, asymmetric fluid pressure force by interrupting the first bearing  68   n  with a first end groove  67  that is in fluid communication with the first port  42  through a first-end-groove connection  88 , thereby exposing the first-piston-rod end surface  136   n  with the pressure at the first port  42 .  
         [0104]     The first-end-groove connection  88  can be functionally replaced, without jeopardizing the radial support for the first piston rod  34   n , by one or more grooves or undercuts (not shown in  FIG. 17 ) on the inner surface of the first bearing  68   n , running longitudinally between the first end groove  67  and the first control bore  110 , and intermittently distributed around the circumference of the first bearing  68   n . If desired, the end snubber valve  208  or the end flow regulator  212  as illustrated in  FIG. 16  can be incorporated to control the end snubber in this embodiment as well.  
         [0105]     In the embodiment in  FIG. 17 , the first and second actuation springs  62  and  58  are pneumatic springs; that is, they include gaseous volumes enclosed in a pneumatic cylinder  254  and separated by a pneumatic piston  250  with an optional pneumatic piston seal  252 . The design of the pneumatic springs can be optionally replaced by other common variations, such as a bladder type of construction (not shown in  FIG. 17 ) for better leakage prevention. The pneumatic cylinder  254  can be fabricated inside the housing  64   n  (as shown in  FIG. 17 ) or in a separate mechanical block.  
         [0106]     The first and second actuation springs  62  and  58  are connected with one or more gas sources (not shown in  FIG. 17 ) through pneumatic first and second ports  260  and  262  respectively and one or more associated pneumatic control valves (not shown in  FIG. 17 ) for leakage compensation, spring stiffness control and optional initialization. Alternatively, it is possible to eliminate one of the pneumatic first and second ports  260  and  262  by allowing a certain leakage between the two pneumatic springs. The spring stiffness control includes regulating and/or changing, in real time per functional needs and operational conditions, the absolute stiffness level and the stiffness differential of the two pneumatic springs. The stiffness differential helps create asymmetric net spring force desired for certain applications. The actuator  30   n  can be initialized by creating a pressure differential across the two springs  62  and  58  at the startup. For example, it can be initialized to a fully closed position by causing higher pressure in the first actuation spring  62  than in the second actuation spring  58 .  
         [0107]     In all the above descriptions, the first and second actuation springs  62  and  58  are each identified or illustrated, for convenience, as a single spring. When needed for strength, durability or packaging, however each or any one of the first and second actuation springs  62  and  58  may include a combination of two or more springs. In the case of mechanical compression springs, they can be nested concentrically, for example. The spring subsystem may also include a single mechanical spring (not shown) that can take both tension and compression. The spring subsystem may also include a combination of pneumatic and mechanical springs.  
         [0108]     Also in many illustrations and descriptions, the fluid medium is assumed to be hydraulic or in liquid form. In most cases, the same concepts can be applied with proper scaling to pneumatic actuators and systems. As such, the term “fluid” as used herein is meant to include both liquids and gases. Also, in many illustrations and descriptions so far, the application of the 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.  
         [0109]     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.

Technology Classification (CPC): 5