Patent Application
20070022988
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.
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.
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 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 adjacent, in the first direction, to and in fluid communication with the first fluid space, whereas the stroke controller contains a second bore adjacent, 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 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, 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 piston rod first end 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.
In further alternative embodiments, either the first-supplemental or second-supplemental flow mechanism may be eliminated by extending the opening range of either the first or second flow mechanism respectively, resulting in simpler and more compact designs.
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.
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;
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;
a is a drawing of different alternative embodiment of the invention, including an end switch valve;
b is a drawing of yet a further alternative embodiment of the invention, including a differently configured end switch valve;
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;
b is a drawing of the same alternative embodiment as in
Referring now to
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
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
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 FM1, FM1S, FM2 and FM2S described in details in
The second flow mechanism FM2 and second-supplemental flow mechanism FM2S, together as a second flow control subsystem, control fluid communication between the second fluid space 86 and the second port 42. The second flow mechanism FM2 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 FM2S 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 FM2 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 FM2S 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 FM1, FM1S, FM2 and FM2S, 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 56 and 42 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 open flow can also be prevented with just one of the two fluid spaces being blocked from its corresponding port, examples of which are some preferred embodiments illustrated later in
It is generally preferred for the first and second necks 39 and 53 to have a circular or cylindrical shape. But when desired it is also feasible for a neck to have an outer dimension substantially smaller than the inner dimension of a corresponding bore only over a portion of the circumference (not shown in the Figures).
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
The continuous control of the stroke for the preferred embodiment shown in
If the function of the continuous or proportional control of the stroke is not needed, the embodiment in
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
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 or position, 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
When the power is off, the status of the system is substantially as that shown in
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 FM1 and the second-supplemental flow mechanism FM2S are slightly open, while the first-supplemental flow mechanism FM1S and the second flow mechanism FM2 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 L1.
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
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 FM1 and the second-supplemental fluid mechanism FM2S, 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
The state in
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 FM1S to the first flow mechanism FM1 for the first fluid space 84 and from the second flow mechanism FM2 to the second-supplemental flow mechanism FM2S 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.
Once the actuation piston 46 disengages or underlaps the first partial cylinder 114, all four flow mechanisms FM1, FM2, FM1S and FM2S, as defined in
Once the actuation piston 46 overlaps or engages the second partial cylinder 115 when the engine valve opening Xev is between (ST−L2) 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 FM1S and the second flow mechanism FM2, 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
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
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.
With the above changes, the first and second-supplemental flow mechanisms FM1S and FM2S in
The second-supplemental flow mechanism FM2S runs between the second port 42 and the second fluid space 86, through the second-supplemental chamber 41, the first rod passage 150e, the second-supplemental chamber extension 112, the E-E-E passage, and the bypass passage 48. The second-supplemental flow mechanism FM2S 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 31e, which may also necessitate lengthening the stroke controller 123e and the housing 64e.
Refer now to
Refer now to
Refer now to
The pins 140 slideably run through pin holes 141 fabricated in the housing 64v. The pin holes 141 are not to interfere with the first and second ports 56 and 42 and associated flow channels as shown in
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 121v, 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 L2 or more in the valve opening stroke, for example, may prevent the actuation piston 46 reaches the second partial cylinder 115 as shown in
Refer now to
With the capped first bore 68w, 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 68w, a potential external leakage site is eliminated.
Refer now to
In
Referring now to
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 68w 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 68w 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
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 68w and, instead of the fourth port 45, the first end groove 67.
The embodiment in
The extra stroke control chamber 222 and the stroke control chamber 125 are more effective in resisting the dynamic motion of the stroke controller 123x 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_ST2 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
The embodiment in
Because of the discontinuous nature of the grooves 108x and 109x around the circumference, some mechanism, such as a tube key 250, is used to prevent the stroke controller 123x from drifting around the circumference and to keep proper alignment and fluid communication between the first groove 108x and the second port 42 and between the second groove 109x 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 64x, in a position as shown in
Refer now to
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_ST2 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 123x. 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
This embodiment further includes a variation in the spatial arrangement of the first and second grooves 108y and 109y, which are relocated from the stroke controller 123y to the housing 64y while still maintaining their functions to keep, regardless the longitudinal position of the stroke controller 123y relative to the actuator housing 64y, uninterrupted fluid communication between the second chamber 104y and the second port 42 and between the first-supplemental chamber 105y and the first port 56, respectively. The grooves 108y and 109y also help keep hydrostatic force balance on the stroke controller 123y. 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
Mathematically, the respective spring forces F1 and F2 from the first and second actuation spring 62y and 58y are
F2=[F2o+K2*(STmax−ST)/2]−K2*(Xev−ST/2) and
F1=−[F1o+K1*(STmax−ST)/2]−K1*(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 F1o and F2o are the respective spring preloads of the first and second actuation spring 62y and 58y 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. K1 and K2 are the respective spring rates. Here the springs 62y and 58y 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
F=[(F2o−F1o)+(K2−K1)*(STmax−ST)/2]−(K2+K1)*(Xev−ST/2)
or
F=Fo−K*(Xev−ST/2),
with Fo are K being the total pre-load and spring rate, and
Fo=(F2o−F1o)+(K2−K1)*(STmax−ST)/2 and
K=K2+K1.
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 STmax, then one has
F2o=F1o,
K1=(K+2*Fmino/STmax)/[2*(1−STmin/STmax)], and
K2=K−K1,
where if K=100,000 N/m, STmin=0.002 m, STmax=0.008 m, and Fmino=20 N, then K1=70,000 N/m and K2=30,000 N/m, i.e., with the first actuation spring rate K1 being substantially higher than the second actuation spring rate K2. Only relative values of the spring preloads F1o and F2o 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, repectively, then one has
(F2o−F1o)−K*(Xev_maxo−STmax/2),
K1=K*(Xev_maxo−Xev_mino)/(STmax−STmin)), and
K2=K−K1.
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>STmin/2 and Xe_maxo>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 K1=K2=50,000 N/m and (F2o−F1o)=50 N, i.e., with the second actuation spring preload F2o being substantially higher than the first actuation spring preload F1o.
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 K1=80,000 N/m, K2=20,000 N/m and (F2o−F1o)=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 62y) and 58 (or 58y) are each identified or illustrated, for convenience, as a single mechanical compression spring. When needed for strength, durability or packaging, each or anyone of the first and second actuation springs 62 or 62y and 58 or 58y may include a combination of two or more mechanical compression springs, nested concentrically for example. The spring subsystem may comprise pneumatic springs (not illustrated), instead of mechanical ones, as long as it is able to exert the actuation piston in both directions and has tendency to bring the actuation piston to a neutral state. The spring subsystem may also include a single mechanical spring (not shown) that can take both tension and compression.
Referring now to
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.
a and 17b 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 FM1 and the first-supplemental flow mechanism FM1S and fluid communication between the second port 42 and the second fluid space 86 through the second-supplemental flow mechanism FM2S and the second flow mechanism FM2. The first-supplemental flow mechanism FM1S is open via the unblocked shuttle valve first bore 274 and the bypass passage 48x as shown in
During the transition from the state in
Dynamics is in a reverse order for the transition from the state in
Relative to the embodiments in
With the first piston rod first end 35x exposed to the pressure at the first port 56, which is under the high pressure P_H during the opening stroke, this arrangement in
Referring now to
With continuing reference to
As with some of the earlier embodiments, one can utilize the closed end of the first bore 68, the first end groove 67, and the first piston rod 34 to provide additional snubbing action when the first piston rod 34 longitudinally overlaps the part of the first bore 68 in the first direction beyond the first end groove 67. A snubbing taper 280 on the rod 34 offers a varying degree of flow restriction. One can optionally utilize the end snubber valve 208 to disable the snubbing function—during non-idle engine operations, for example—by switching the end snubber valve to its left or open position and short-circuiting the first end groove 67 and the closed end of the first bore 68. The end snubber valve 208 illustrated in
The closed end of the first bore 68 and the first end groove 67 are supplied through the fourth port 45 by a fluid supply at a pressure of P_END, the value or level of which can be selected per functional needs. The supply can be, for example, fixed at the low system pressure P_L for simple snubbing function. Alternatively, it can be equal to the pressure at the first port 56, which alternates between the high and low system pressures P_H and P_L. This creates a flow passage, not shown in
As a design alternative, the second flow mechanism FM2 may include one or more fluid passages (not shown in
As with the embodiments illustrated in
To achieve a maximum opening force according to the embodiment illustrated in
Again for simplification and emphasis on variations,
The actuation switch valve 80 in
One can purposely introduce a time delay between the actions of the two actuation switch valves 80a and 80b 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 80a 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.
This application is a continuation-in-part of U.S. patent application Ser. No. 11/292,879, filed Dec. 2, 2005, which is a continuation-in-part of U.S. patent application Ser. No. 11/194,243, filed Aug. 1, 2005, the entire content of both of which are incorporated herein by reference.
| Number | Date | Country | |
|---|---|---|---|
| Parent | 11292879 | Dec 2005 | US |
| Child | 11325986 | Jan 2006 | US |
| Parent | 11194243 | Aug 2005 | US |
| Child | 11292879 | Dec 2005 | US |
