Patent Publication Number: US-2016222841-A1

Title: Variable lost motion valve actuator and method

Description:
CROSS REFERENCE TO RELATED PATENT APPLICATIONS 
     This application is a continuation of and claims the benefit of priority to U.S. patent application Ser. No. 14/323,385, filed Jul. 3, 2014, which in turn is a continuation of and claims the benefit of priority to U.S. patent application Ser. No. 14/139,308, filed Dec. 23, 2013, and issued as U.S. Pat. No. 8,776,738, which in turn is a continuation of and claims the benefit of priority to U.S. patent application Ser. No. 13/021,531, filed Feb. 4, 2011, and issued as U.S. Pat. No. 8,820,276, which in turn is a continuation of and claims the benefit of priority to U.S. patent application Ser. No. 11/450,286, filed Jun. 12, 2006 and issued as U.S. Pat. No. 7,882,810, which in turn is a continuation-in-part of and claims the benefit of priority to U.S. patent application Ser. No. 10/251,748, filed Sep. 23, 2002 and issued as U.S. Pat. No. 7,059,282, which in turn is a divisional of and claims the benefit of priority to U.S. patent application Ser. No. 09/749,907, filed Dec. 29, 2000 and issued as U.S. Pat. No. 6,510,824, which in turn is a continuation-in-part of and claims the benefit of priority to U.S. patent application Ser. No. 09/594,791, filed Jun. 16, 2000 and issued as U.S. Pat. No. 6,293,237, which in turn is a continuation of and claims the benefit of priority to U.S. patent application Ser. No. 09/209,486, filed Dec. 11, 1998 and issued as U.S. Pat. No. 6,085,705, which in turn claims the benefit of priority to U.S. Provisional Application Ser. No. 60/069,270, filed Dec. 11, 1997. 
    
    
     FIELD OF THE INVENTION 
     The present invention relates generally to methods and apparatus for intake and exhaust valve actuation in internal combustion engines. 
     BACKGROUND OF THE INVENTION 
     Valve actuation in an internal combustion engine is required in order for the engine to produce positive power, as well as to produce engine braking. During positive power, intake valves may be opened to admit fuel and air into a cylinder for combustion. The exhaust valves may be opened to allow combustion gas to escape from the cylinder. 
     During engine braking, the exhaust valves may be selectively opened to convert, at least temporarily, an internal combustion engine into an air compressor. This air compressor effect may be accomplished by partially opening one or more exhaust valves near piston top dead center position for compression-release type braking, or by maintaining one or more exhaust valves in a partially open position for much or all of the piston motion for bleeder type braking. In doing so, the engine develops retarding horsepower to help slow the vehicle down. This can provide the operator increased control over the vehicle and substantially reduce wear on the service brakes of the vehicle. A properly designed and adjusted engine brake can develop retarding horsepower that is a substantial portion of the operating horsepower developed by the engine in positive power. 
     The braking power of an engine brake may be increased by selectively opening the exhaust and/or intake valves to carry out exhaust gas recirculation (EGR) in combination with engine braking. Exhaust gas recirculation denotes the process of channeling exhaust gas back into the engine cylinder after it is exhausted out of the cylinder. The recirculation may take place through the intake valve or the exhaust valve. When the exhaust valve is used, for example, the exhaust valve may be opened briefly near bottom dead center on the intake stroke of the piston. Opening of the exhaust valve at this time permits higher pressure exhaust gas from the exhaust manifold to recirculate back into the cylinder. The recirculation of exhaust gas increases the total gas mass in the cylinder at the time of the subsequent engine braking event, thereby increasing the braking effect realized. 
     For both positive power and engine braking applications, the engine cylinder intake and exhaust valves may be opened and closed by fixed profile cams in the engine, and more specifically by one or more fixed lobes which may be an integral part of each of the cams. The use of fixed profile cams makes it difficult to adjust the timings and/or amounts of engine valve lift needed to optimize valve opening times and lift for various engine operating conditions, such as different engine speeds. 
     One method of adjusting valve timing and lift, given a fixed cam profile, has been to incorporate a “lost motion” device in the valve train linkage between the valve and the cam. Lost motion is the term applied to a class of technical solutions for modifying the valve motion dictated by a cam profile with a variable length mechanical, hydraulic, or other linkage means. In a variable valve actuation lost motion system, a cam lobe may provide the “maximum” (longest dwell and greatest lift) motion needed for a full range of engine operating conditions. A variable length system may then be included in the valve train linkage, intermediate of the valve to be opened and the cam providing the maximum motion, to subtract or lose part or all of the motion imparted by the cam to the valve. 
     This variable length system (or lost motion system) may, when expanded fully, transmit all of the cam motion to the valve, and when contracted fully, transmit none or a partial amount of the cam motion to the valve. An example of such a system and method is provided in Vorih et al., U.S. Pat. No. 5,829,397 (Nov. 3, 1998), Hu, U.S. Pat. No. 6,125,828, and Hu U.S. Pat. No. 5,537,976, which are assigned to the same assignee as the present application, and which are incorporated herein by reference. 
     In some lost motion systems, an engine cam shaft may actuate a master piston which displaces fluid from its hydraulic chamber into a hydraulic chamber of a slave piston. The slave piston in turn acts on the engine valve to open it. The lost motion system may include a solenoid valve and a check valve in communication with a hydraulic circuit connected to the chambers of the master and slave pistons. The solenoid valve may be maintained in an open or closed position in order to retain hydraulic fluid in the circuit. As long as the hydraulic fluid is retained, the slave piston and the engine valve respond directly to the motion of the master piston, which in turn displaces hydraulic fluid in direct response to the motion of a cam. When the solenoid position is changed temporarily, the circuit may partially drain, and part or all of the hydraulic pressure generated by the master piston may be absorbed by the circuit rather than be applied to displace the slave piston. 
     The complexity of, and challenges posed by, lost motion systems may be increased by the need to incorporate an adequate fail-safe or “limp home” capability into such systems. In previous lost motion systems, a leaky hydraulic circuit could disable the master piston&#39;s ability to open its associated valve(s). If a large enough number of valves cannot be opened at all, the engine cannot be operated. Therefore, one valuable feature of various embodiments of the invention arises from the ability to provide a lost motion system which enables the engine to operate at some minimum level (i.e. at a limp home level) should the hydraulic circuit of such a system develop a leak. A limp home mode of operation may be provided by using a lost motion system which still transmits a portion of the cam motion to the valve after the hydraulic circuit associated with the cam leaks or the control thereof is lost. In this manner the most extreme portions of a cam profile still can be used to get some valve actuation after control over the variable length of the lost motion system is lost and the system has contracted to a reduced length. The foregoing assumes, of course, that the lost motion system is constructed such that it will assume a contracted position should control over it be lost and that the valve train will provide the valve actuation necessary to operate the engine. In this manner the lost motion system may be designed to allow the engine to operate such that an operator can still “limp home” and make repairs. 
     A fundamental feature of lost motion systems is their ability to vary the length of the valve train. Not many lost motion systems, however, have utilized the high speed mechanisms that are required to rapidly vary the length of the lost motion system on a valve event-by-event basis. Lost motion systems accordingly have not been variable such that they may assume two functional lengths per cycle of the engine. The lost motion system that is the subject of this application is considerably advanced in comparison to other known systems due to its ability to provide variable valve actuation (VVA) on a valve event-by-event basis with each cycle of the engine. By using a high speed mechanism to vary the length of the lost motion system, more precise control may be attained over valve actuation, and accordingly optimal valve actuation may be attained for a wide range of engine operating conditions. 
     Applicants have determined that the lost motion system and method of the present invention may be particularly useful in engines requiring valve actuation for positive power, compression release engine braking, exhaust gas recirculation, cylinder flushing, and low speed torque increase. Typically, compression release and exhaust gas recirculation events involve much less valve lift than do positive-power-related valve events. Compression release and exhaust gas recirculation events may, however, require very high pressures and temperatures to occur in the engine. Accordingly, if left uncontrolled (which may occur with the failure of a lost motion system), compression release and exhaust gas recirculation could result in pressure or temperature damage to an engine at higher operating speeds. Therefore, it may be beneficial to have a lost motion system which is capable of providing control over positive power, compression release, and exhaust gas recirculation events, and which will provide only positive power or some low level of compression release and exhaust gas recirculation valve events, should the lost motion system fail. It may also be beneficial to provide a lost motion system capable of providing post main exhaust valve events which may be used to achieve cylinder flushing and low speed torque increases. 
     An example of a lost motion system and method used to obtain retarding and exhaust gas recirculation is provided by the Gobert, U.S. Pat. No. 5,146,890 (Sep. 15, 1992) for a Method And A Device For Engine Braking A Four Stroke Internal Combustion Engine, assigned to AB Volvo, and incorporated herein by reference. Gobert discloses a method of conducting exhaust gas recirculation by placing the cylinder in communication with the exhaust system during the first part of the compression stroke and optionally also during the latter part of the inlet stroke. Gobert uses a lost motion system to enable and disable retarding and exhaust gas recirculation, but such system is not variable within an engine cycle. 
     In view of the foregoing, there is a significant need for a system and method of controlling lost motion which: (i) optimizes engine operation under various engine operating conditions; (ii) provides precise control of lost motion; (iii) provides acceptable limp home and engine start-up capability; and (iv) provides for high speed variation of the length of a lost motion system. The lost motion system that is the subject of this application meets these needs, as well as others. 
     As noted above, one constraint on the use of lost motion systems arises from the addition of bulk in the engine compartment. Known systems for providing lost motion valve actuation have tended to be non-integrated devices which add considerable bulk to the valve train. As vehicle dimensions have decreased, so have engine compartment sizes. Accordingly, there is a need for a less bulky lost motion system, and in particular for a system which is compact and has a relatively low profile. 
     Furthermore, there is a need for low profile lost motion systems capable of varying valve actuation responsive to engine and ambient conditions. Variable actuation of intake and exhaust valves in an internal combustion engine may be useful for all potential valve events (positive power and engine braking). When the engine is in positive power mode, variation of the opening and closing times of intake and exhaust valves may be used in an attempt to optimize fuel efficiency, power, exhaust cleanliness, exhaust noise, etc., for particular engine and ambient conditions. During engine braking, variable valve actuation may enhance braking power and decrease engine stress and noise by modifying valve actuation as a function of engine and ambient conditions. 
     In an attempt to develop a functional and robust variable valve actuation system that is useful for both positive power and engine braking applications, Applicants have had to solve several design challenges. These design challenges have resulted in the development of sub-systems that not only allow the subject system to work effectively, but which may also be useful in other variable valve actuation systems. 
     For example, engine valves are required to open and close very quickly, therefore the valve spring is typically very stiff. When the valve closes, it may impact the valve seat with such force that it eventually erodes the valve or the valve seat, or even cracks or breaks the valve. In mechanical valve actuation systems that use a valve lifter to follow a cam profile, the cam lobe shape provides built-in valve-closing velocity control. In common rail hydraulically actuated valve assemblies, however, there is no cam to self-dampen the closing velocity of an engine valve. Likewise, in hydraulic lost motion systems such as the present ones, a rapid draining of fluid from the hydraulic circuit may allow an engine valve to “free fall” and seat at an unacceptably high velocity. 
     The system that is the subject of this application, being a lost motion system, presents valve seating challenges. The variable valve actuation capability of the present system may result in the closing of an engine valve at an earlier time than that provided by the cam profile. This earlier closing may be carried out by rapidly releasing hydraulic fluid (to an accumulator in the preferred embodiment) in the lost motion system. In such instances engine valve seating control is required because the rate of closing the valve is governed by the hydraulic flow to the accumulator instead of by the fixed cam profile. Engine valve seating control may also be required for applications (e.g. centered lift) in which the engine valve seating occurs on a high velocity region of the cam. 
     Applicants approached the valve seating challenge with the understanding that valve seating velocity should be less than approximately 0.4 m/sec. Absent steps to control valve seating velocity, however, the valves could seat at a velocity that is an order of magnitude greater. Applicants also determined that valve seating control preferably should be designed to function when the closing valve gets within 0.5 to 0.75 mm of the valve seat. The combination of valve thermal growth, valve wear, and tolerance stack-up can exceed 0.75 mm, resulting in the complete absence of seating velocity control or in an exceedingly long seating event if measures are not taken to adjust the lash of the valve seating control to account for such variations. It is also assumed that, preferably, valve seating control should not significantly reduce initial engine valve opening rate, and valve seating control should be capable of operating over a wide range of valve closing velocities and oil viscosities. 
     Existing devices used to control valve seating velocity may use hydraulic fluid flow restriction to produce pressure that acts on an area of the slave piston to develop a force to slow the slave piston and reduce seating velocity. The area on which the pressure acts may be very small in such devices which in turn requires that the pressure opposing the valve return spring be high, and the controlling flow rate be low. Low controlling flow rates result in an increased sensitivity to leakage. In addition, these devices may restrict the hydraulic fluid flow that produces valve opening. 
     In view of the foregoing there is a need for a valve catch sub-system for valve seating control that provides fine control of hydraulic fluid flow through the sub-system. There is also a need for a sub-system that does not adversely affect hydraulic fluid flow for valve opening and which is less susceptible to dimensional tolerances affecting leakage. In particular, there is a need for valve seating that is improved by a flow control that becomes more restrictive as the valve approaches the seat. 
     There is also a need for a valve catch that adjusts for lash differences between the engine valve and the valve catch. Although most variable valve actuation (VVA) systems are inherently self lash adjusting, valve seating control is not. Systems that do not need manual adjustment, either initially or as the system ages, are desirable. Previous valve seating control mechanisms have required a manual lash adjustment or a separate set of lash adjustment hardware The design of a conventional hydraulic lash adjustor capable of transmitting compression-release braking loads would be challenging due to structural and compliance requirements. 
     The valve catch embodiment(s) of the present invention meet the aforementioned needs and provide other benefits as well. The valve catch embodiment(s) disclosed herein provide acceptable engine valve seating velocity in a VVA system, such as a lost motion or common rail system. For a lost motion VVA system, engine valve seating control is provided for early engine valve closing, where the rate of closing is governed by the hydraulic flow from the control piston to the accumulator as opposed to a cam profile. Engine valve seating control also may be provided for a high velocity region of the cam. The lash adjusting portion of this mechanism provides an additional amount of seating control for the last few hundredths of a millimeter of valve closing. 
     The valve catch embodiment(s) of the present invention includes a variable flow area in the sub-system plunger. The valve catch embodiment(s) of the invention may also be designed to have relatively high flow rates, large orifices, and utilize small pressure drops. The valve catch embodiment(s) of the present invention may also experience reduced peak valve catch pressure as compared with some known valve catch systems. Furthermore, the variable flow restriction design of the valve catch embodiment(s) of the present invention is expected to be more robust than the constant flow restriction design with respect to engine valve velocity at the point of valve catch engagement and oil temperature and aeration control. Variable flow restriction may allow the displacement at the point of valve catch/slave piston engagement to be reduced, so that the valve catch has less undesired effect on the breathing of the engine. 
     Furthermore, Applicants implementation of a variable valve actuation system using lost motion hydraulic principles may require a sub-system for effecting initial start up of the system. An initial start mechanism (ISM) may be required to (i) accelerate the process of charging the subject lost motion system with hydraulic fluid, and/or (ii) permit actuation of the engine valve until such time as the subject system is fully charged with hydraulic fluid. Absent such a system, starting and/or smooth operation of the engine could be delayed due to the inaction of the engine valves until there is sufficient hydraulic fluid in the system to produce the desired valve motions. An added advantage of such a system is that it may provide a limp-home mode of operation for the engine as well in the event that the system is incapable of being charged with hydraulic fluid. Therefore, there is a need for a sub-system that provides valve actuation between the initial cranking of an engine and the charging of the variable valve actuation system with hydraulic fluid. 
     Still other advancements that may be required for operation of the subject system include an accumulator sub-system. In order to broaden the range of possible valve actuations that may be produced with the subject system, it may be beneficial to improve the rate at which the accumulator can absorb fluid and the rate at which it can supply fluid for re-fill operations. Improvement of this response time may permit more rapid variation of the motion of the engine valves in the system and may limit the loss of cam follow during periods of hydraulic fluid flow from the accumulator to the high-pressure hydraulic circuit. Accordingly, there is a need for a system accumulator with improved response time. 
     A basic method of improving accumulator response time is to increase the strength of the spring biasing the accumulator piston into its refill position. However, accumulator spring force cannot be increased indefinitely without incurring associated costs. For example, the accumulator spring force should be limited relative to the engine valve spring force so as to avoid engine valve float. In turn, the engine valve spring force may be limited by spring envelope constraints and the need to minimize parasitic loss of the VVA system. 
     Furthermore, the accumulator design would ideally prevent the high-pressure circuit pressure from dropping below ambient or the accumulator piston from bottoming out in its bore, because these situations could cause cavitation and evolution of dissolved air in the oil. This problem may be particularly troublesome during an early engine valve closing event, where oil must quickly flow to the accumulator to affect the early closing and then flow back to the high-pressure circuit when the engine valve seats or valve catch engages. 
     Despite all of the foregoing design challenges, Applicants have designed a compact and efficient accumulator system that provides improved response time. Applicants have designed a relatively low pressure accumulator system which provides improved performance as the result of synergy attributable to the combination of a low restriction trigger valve, shorter and larger fluid passages between the system elements, use of fewer or no check valves, larger yet low inertia accumulator pistons, reduced accumulator piston travel, and a gallery arrangement of multiple accumulators in common hydraulic communication. 
     Control feature advancements also appear to be desirable in view of the capabilities of the subject VVA system. For example, in some embodiments of the present invention, each of the engine valves in the subject system may be independently turned “on” or “off’ for a prolonged period. Accordingly, there is a need for advanced control features, such as cylinder cut-out capability, which may reduce fuel consumption by only activating individual engine valves or engine valves associated with individual cylinders, on an as needed basis. 
     Control over cylinder cut-out necessarily requires active control over cylinder re-start. Assuming the cylinder cut-out is controlled in response to engine load (the lower the load, the less cylinders needed for power), then cylinder re-start must also be provided responsive to increasing engine load. Embodiments of the present invention provide for such active control over cylinder re-start, as well as cylinder cut-out. 
     The use of hydraulic actuation also may necessitate control features that modify the timing of hydraulic actuation based on the viscosity of the hydraulic fluid in the system. Typically, the viscosity of hydraulic fluid, such as engine oil, lowers as it increases in temperature. As viscosity lowers, the response time for hydraulic actuation involving the fluid may decrease. Because the temperature of the hydraulic fluid used in connection with the various embodiments of the present invention may vary by more than 100 degrees Celsius, there is a need to adjust the timing of some hydraulic actuation events based on the temperature and/or viscosity of the hydraulic fluid. Various embodiments of the present invention provide for modification of hydraulic actuation based on the temperature and/or viscosity of the hydraulic fluid used for such actuation. 
     Others have attempted to provide for the modification of valve actuation systems. U.S. Pat. No. 5,423,302 to Glassey discloses a fuel injection control system having actuating fluid viscosity feedback using several sensors including a crankshaft angular speed sensor, an engine coolant temperature sensor, and a voltage sensor. U.S. Pat. No. 5,411,003 to Eberhard et al. (“Eberhard”) discloses a viscosity sensitive auxiliary circuit for a hydromechanical control valve for timing the control of a tappet system. Eberhard utilizes a pressure divider chamber to influence timing control. U.S. Pat. No. 4,889,085 to Yagi et al. discloses a valve operating device for an internal combustion engine that utilizes a damper chamber in connection with a restriction mechanism. Some of these inventions attempt to compensate for increased viscosity by modifying the flow of working fluid, rather than the timing of the operation of the valves themselves. In addition, many of these devices are complex and difficult to maintain. Accordingly, there remains a need for a method and apparatus for modifying the opening and closing of engine valves based on an engine fluid temperature and/or viscosity that is accurate, easy to implement, cost effective, and easy to calibrate by the user. 
     As may be evident, the embodiments of the present invention disclosed herein may be particularly useful in a wide variety of internal combustion engines. Such engines are often considered to emit undesirably high levels of noise. Accordingly, various embodiments of the invention may also incorporate control features which tend to reduce the level of noise produced by such engines, both during positive power and during engine braking. 
     OBJECTS OF THE INVENTION 
     It is therefore an object of the present invention to provide a system and method for optimizing engine operation under various engine and ambient operating conditions through variable valve actuation control. 
     It is another object of the present invention to provide a system and method for providing high speed control of the lost motion in a valve train. 
     It is a further object of the present invention to provide a system and method of valve actuation which provides a limp-home capability. 
     It is yet another object of the present invention to provide a system and method for selectively actuating a valve with a lost motion system for positive power, compression release braking, and exhaust gas recirculation modes of operation. 
     It is still a further object of the present invention to provide a system and method for valve actuation which is compact and light weight. 
     It is still another object of the present invention to provide a system and method for seating an engine valve after actuation thereof. 
     It is still another object of the present invention to provide a system and method for actuating the engine valves in a lost motion system prior to charging the system with hydraulic fluid. 
     It is still another object of the present invention to provide a system and method for accelerating the process of charging a lost motion system with hydraulic fluid. 
     It is still another object of the present invention to provide a system and method for improving the response time of the accumulator used in a variable valve actuation system. 
     It is still another object of the present invention to provide a system and method for selectively cutting-out and re-starting the operation of engine valves for particular cylinders. 
     It is still another object of the present invention to provide a system and method for improving positive power fuel economy of an engine. 
     It is still another object of the present invention to provide a system and method for decreasing the noise produced by an engine, particularly compression release engine braking noise. 
     It is still another object of the present invention to provide a system and method for decreasing emissions produced by an engine. 
     It is still another object of the present invention to provide a system and method for modifying the timing of hydraulic actuation in a variable valve actuation system to account for changes in hydraulic fluid temperature and/or viscosity. 
     It is still another object of the present invention to provide systems and methods for hydraulically and electronically controlling the actuation of engine valves for positive power and engine braking applications. 
     Additional objects and advantages of the invention are set forth, in part, in the description which follows, and, in part, will be apparent to one of ordinary skill in the art from the description and/or from the practice of the invention. 
     SUMMARY OF THE INVENTION 
     In response to this challenge, Applicants have developed an innovative and reliable engine valve actuation system comprising: means for containing the system; a piston bore provided in the system containing means; a low pressure fluid supply passage connected to the piston bore; a piston having (i) a lower end residing in the piston bore, and (ii) an upper end extending out of the piston bore; a pivoting lever including first, second, and third contact points, wherein the first contact point of the lever is adapted to impart motion to the engine valve, and the third contact point is adapted to contact the piston upper end; a motion imparting valve train element contacting the second contact point of the pivoting lever; and means for repositioning the piston relative to the piston bore, said means for repositioning intersecting the low pressure fluid supply passage. 
     Applicants have also developed an innovative engine valve actuation system adapted to selectively provide main valve event actuations and auxiliary valve event actuations, said system comprising: means for containing the system, said containing means having a piston bore and a first fluid passage communicating with the piston bore; a lever located adjacent to the containing means, said lever including (i) a first repositionable end, (ii) a second end for transmitting motion to an engine valve, and (iii) a centrally located cam roller; a piston disposed in the piston bore and connected to the first repositionable end of the lever; a cam in contact with the cam roller; a fluid control valve in communication with the piston bore via the first fluid passage; means for actuating the fluid control valve to control the flow of fluid from the piston bore through the first fluid passage; and means for supplying low pressure fluid to the piston bore. 
     Applicants have further developed an innovative apparatus for limiting the seating velocity of an engine valve comprising: a housing; a seating bore provided in the housing; means for supplying fluid to the seating bore; an outer sleeve slidably disposed in the seating bore and defining an interior chamber; a cup piston slidably disposed in the outer sleeve, said cup piston having a lower surface adapted to transmit a valve seating force to the engine valve; a cap connected to an upper portion of the outer sleeve, said cap having an opening there through; a disk disposed within the interior chamber between the cup piston and the cap, said disk having at least one opening there through; a central pin disposed in the interior chamber between the cup piston and the disk; a spring disposed around the central pin and between the disk and the cup piston; an upper seating member slidably disposed in the seating bore; and a means for biasing the upper seating member towards the cap. 
     Applicants have also developed an innovative valve actuation system for controlling the operation of an engine valve, said system comprising: means for hydraulically varying the amount of engine valve actuation; a solenoid actuated trigger valve operatively connected to the means for hydraulically varying; and means for determining trigger valve actuation and deactuation times based on a selected engine mode, and engine load and engine speed values. 
     Applicants have further developed an innovative valve actuation system for controlling the operation of at least one valve of an engine at different operating temperatures, comprising: means for determining a present temperature of an engine fluid; means for operating the at least one valve; and means for modifying the operation of the at least one valve in response to the determined temperature. 
     Applicants have also developed an innovative valve actuation system for controlling the operation of at least one valve of an engine at different engine fluid operating viscosities, comprising: means for determining a present viscosity of an engine fluid; means for operating the at least one valve; and means for modifying the operation of the at least one valve in response to the determined viscosity. 
     Applicants have further developed an innovative method of modifying the timing of at least one engine valve, said method comprising the steps of: determining a current temperature of an engine fluid; determining a timing modification for the operation of the at least one engine valve based on the determined current temperature; and modifying the timing of the operation of the at least one engine valve in response to the determined timing modification. 
     Applicants have also developed an innovative method of modifying the timing of at least one engine valve, said method comprising the steps of: determining a current viscosity of an engine fluid; determining a timing modification for the operation of the at least one engine valve based on the determined current viscosity; and modifying the timing of the operation of the at least one engine valve in response to the determined timing modification. 
     Applicants have further developed an innovative lost motion engine valve actuation system comprising: a rocker lever adapted to provide engine valve actuation motion, said rocker lever having a first repositionable end and a second end for transmitting valve actuation motion; means for hydraulically varying the position of the first end of the rocker lever; and means for maintaining the position of the first end of the rocker lever during periods of time that the means for hydraulically varying is inoperative. 
     Applicants have still further developed an innovative lost motion engine valve actuation system comprising: a lost motion piston and a means for locking said lost motion piston into a fixed position during engine start-up. 
     Applicants have further developed an innovative lost motion engine valve actuation system comprising: a lost motion piston and a means for locking said lost motion piston into a fixed position at times when hydraulic fluid pressure is below a predetermined threshold. 
     It is to be understood that both the foregoing general description and the following detailed description are exemplary and explanatory only, and are not restrictive of the invention as claimed. The accompanying drawings, which are incorporated herein by reference, and which constitute a part of this specification, illustrate certain embodiments of the invention and, together with the detailed description, serve to explain the principles of the present invention. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       Various embodiments and elements of the invention are shown in the following figures, in which like reference numerals are intended to refer to like elements. 
         FIG. 1  is a cross-section of a variable valve actuation system embodiment of the invention. 
         FIG. 2  is a pictorial illustration of a pivoting bridge element of the present invention. 
         FIG. 3  is a pictorial illustration of an alternative pivoting bridge element of the present invention. 
         FIG. 4  is a cross-section of an alternative variable valve actuation system embodiment of the invention. 
         FIG. 5  is a pictorial illustration of an alternative pivoting bridge element of the present invention. 
         FIG. 6  is a cross-section of a second variable valve actuation system embodiment of the invention. 
         FIG. 6A  is a cross-section of the variable valve actuation system shown in  FIG. 6  with the addition of an optional bypass passage connecting the first passage  326  and the second passage  346 . 
         FIG. 7  is a cross-section of an embodiment of the trigger valve portion of the present invention. 
         FIG. 8  is a side view of an embodiment of the valve stem contact pin portion of the present invention. 
         FIG. 9  is a pictorial view of an embodiment of the y-bridge lever portion of the present invention. 
         FIG. 10  is a cross-section of an embodiment of the valve catch portion of the present invention. 
         FIGS. 11, 12, 14, 16, and 18  are top plan views of various embodiments of the rocker lever portion of the present invention. 
         FIG. 13  is a cross-section of a third variable valve actuation system embodiment of the invention. 
         FIG. 15  is a cross-section of a fourth variable valve actuation system embodiment of the invention. 
         FIG. 17  is a cross-section of a fifth variable valve actuation system embodiment of the invention. 
         FIG. 19  is a cross-section of a sixth variable valve actuation system embodiment of the invention. 
         FIG. 20  is a cross-section of a first embodiment of the ISM portion of the present invention. 
         FIG. 21  is a cross-section of a second embodiment of the ISM portion of the present invention. 
         FIGS. 22 and 24  are cross-sections of a third embodiment of the ISM portion of the present invention. 
         FIG. 23  is a cross-section of a fourth embodiment of the ISM portion of the present invention. 
         FIG. 25  is a cross-section of a fifth embodiment of the ISM portion of the present invention. 
         FIG. 26  is a pictorial view of a sixth embodiment of the ISM portion of the present invention. 
         FIG. 27  is a cross-section of a seventh embodiment of the ISM portion of the present invention. 
         FIG. 28  is a pictorial view of a sliding member used in the seventh embodiment of the ISM portion of the present invention shown in  FIG. 27 . 
         FIG. 29  is a pictorial view of an eighth embodiment of the ISM portion of the present invention. 
         FIG. 30  is an elevational view of a ninth embodiment of the ISM portion of the present invention. 
         FIG. 31  is a cut-away pictorial view of a tenth embodiment of the ISM portion of the present invention. 
         FIG. 32  is a cross-section of an eleventh embodiment of the ISM portion of the present invention. 
         FIG. 33  is a cross-section of a twelfth embodiment of the ISM portion of the present invention. 
         FIGS. 34-37  are top plan and side views of a thirteenth embodiment of the ISM portion of the present invention. 
         FIGS. 38-40  are a top plan and cross-section views of a fourteenth embodiment of the ISM portion of the present invention. 
         FIG. 41  is a cross-section of a fifteenth embodiment of the ISM portion of the present invention. 
         FIG. 42  is a schematic diagram of an hydraulic fluid supply system embodiment for use in the present invention. 
         FIG. 43  is a cross-section of a second hydraulic fluid supply system embodiment for use in the present invention. 
         FIG. 44  is a cross-section of an alternative plunger locking device for use in the hydraulic fluid supply system shown in  FIG. 43 . 
         FIG. 45  is a cross-section of an embodiment of a low pressure accumulator for use in the present invention. 
         FIG. 46  is a cross-section of a third hydraulic fluid supply system embodiment for use in the present invention. 
         FIG. 47  is a cross-section of a fourth hydraulic fluid supply system embodiment for use in the present invention. 
         FIG. 48  is a cross-section of a fifth hydraulic fluid supply system embodiment for use in the present invention. 
         FIG. 49  is a cross-section of an sixth hydraulic fluid supply system embodiment for use in the present invention. 
         FIG. 50  is a cross-section of a seventh hydraulic fluid supply system embodiment for use in the present invention. 
         FIG. 51  is a cross-section of an eighth hydraulic fluid supply system embodiment for use in the present invention. 
         FIG. 52  is a cross-section of a ninth hydraulic fluid supply system embodiment for use in the present invention. 
         FIG. 53  is a schematic diagram of an embodiment of an accumulator system for use in the present invention. 
         FIG. 54  is a cross-section of an embodiment of a high pressure accumulator for use in an alternative embodiment of the present invention. 
         FIG. 55  is a bottom plan view of the accumulator piston shown in  FIG. 54 . 
         FIG. 56  is a top plan view of the accumulator piston shown in  FIG. 54 . 
         FIG. 57  is a cross-section of an alternative embodiment of a high pressure accumulator that may be used in the present invention. 
         FIG. 58  is a detailed cross-section of the sealing arrangement shown in  FIG. 57 , showing a de-aeration element and a housing boss. 
         FIG. 59  is a block diagram of the various engine modes used by the electronic valve controller, and the relationship of the modes to each other. 
         FIG. 60  is a pictorial representation of a valve timing map set used to control valve actuation during particular engine operating modes. 
         FIGS. 61-69  are flow charts illustrating various engine control algorithms used for cylinder cut-out and cylinder re-start. 
         FIGS. 70-72  are flow charts illustrating various engine control algorithms used to effect quiet mode engine braking operation. 
         FIGS. 73-75  are graphs used to illustrate the effect of exhaust valve braking event timing on engine braking noise level. 
         FIG. 76  is a flow chart illustrating an algorithm for controlling the operation of at least one engine valve in response to measured or calculated temperature information. 
         FIG. 77  is a flow chart illustrating an algorithm for controlling the operation of at least one engine valve in response to measured or calculated viscosity information. 
         FIG. 78  is a flow chart illustrating an algorithm for controlling the operation of at least one engine valve in response to sensed changes in hydraulic fluid viscosity. 
         FIGS. 79-80  are graphs illustrating the effect of modifying the opening and closing of an electro-hydraulic valve in response to temperature. 
         FIGS. 81-83  are partial cross-sections of valve actuation systems including alternative hydraulic piston locking devices, preferably for ISM. 
     
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     Reference will now be made in detail to a first embodiment of the present invention, an example of which is illustrated in the accompanying drawings. A first embodiment of the present invention is shown in  FIG. 1  as an engine valve actuation system  10 . 
     Engine valve actuation system  10  may include a means for providing valve actuation motion  100 . The motion means  100  may include various valve train elements, such as a cam  110 , a cam roller  120 , a rocker arm  130 , and a lever pushrod  140 . A fixed valve actuation motion may be provided to the motion means  100  via one or more lobes  112  on the cam  110 . Displacement of the roller  120  by the cam lobe  112  may cause the rocker arm  130  to pivot about an axle  132 . Pivoting of the rocker arm  130  may, in turn, cause the lever pushrod  140  to be displaced linearly. The particular arrangement of elements that comprise the motion means  100  may not be critical to the invention. For example, cam  110  alone could provide the linear displacement provided by the combination of cam  110 , roller  120 , rocker arm  130 , and lever pushrod  140 , in  FIG. 1 . 
     Motion means  100  may contact a pivoting bridge  200  at a pivot point  211  (which may or may not be recessed in the bridge). The position of the surface  220  may be adjusted by adjusting the position of the surface on which the surface  220  rests. The pivoting bridge  200  may also include a surface  220  for contacting an adjustable piston  320 , and a surface  230  for contacting a valve stem  400 . Valve springs (not shown) may bias the valve stem  400  upward and cause the surface  220  to be biased downward against a system  300  for providing a moveable surface. 
     System  300  may include a housing  310 , a piston  320 , a trigger valve  330 , and an accumulator  340 . The housing  310  may include multiple passages therein for the transfer of hydraulic fluid through the system  300 . A first passage  326  in the housing  310  may connect the bore  324  with the trigger valve  330 . A second passage  346  may connect the trigger valve  330  with the accumulator  340 . A third passage  348  (see, e.g.,  FIG. 6 ) may connect the accumulator  340  with a check valve  350 . 
     The piston  320  may be slidably disposed in a piston bore  324  and biased upward against the surface  220  by a piston spring  322 . The biasing force provided by the piston spring  322  may be sufficient to hold the piston  320  against the surface  220 , but not sufficient to resist the downward displacement of the piston when a significant downward force is applied to the piston by the surface  220 . 
     The accumulator  340  may include an accumulator piston  341  slidably disposed in an accumulator bore  344  and biased downward by an accumulator spring  342 . Hydraulic fluid that passes through the trigger valve  330  may be stored in the accumulator  340  until it is used to refill the bore  324 . 
     Linear displacement may be provided by the motion means  100  to the pivoting bridge  200 . Displacement provided to the pivoting bridge  200  may be transmitted through surface  230  to the valve stern  400 . The valve actuation motion that is transmitted by the pivoting bridge  200  to the valve stem  400  may be controlled by controlling the position of the surface  220  relative to the pivot point  211 . Given the input of a fixed downward motion on the pivoting bridge  200  by the pushrod  140 , if the position of the surface  220  is raised relative to the pivot point  211 , then the downward motion experienced by the valve stem  400  is increased relative to what it would have otherwise been. Conversely, if the position of the surface  220  is lowered relative to the pivot point  211 , then the downward motion experienced by the valve stem  400  is decreased. Thus, by selectively lowering the position of the surface  220 , relative to the pivot point  211 , motion imparted by the motion means  100  to the pivoting bridge  200  may be selectively “lost”. 
     When the motion means  100  applies a downward displacement to the pivoting bridge  200 , the displacement experienced by the valve stem  400  may be controlled by controlling the position of piston  320  at the time of such downward displacement. During such downward displacement, piston  320  pressurizes the hydraulic fluid in bore  324  beneath the piston. The hydraulic pressure is transferred by the fluid through passage  326  to the trigger valve  330 . Thus, selective bleeding of hydraulic fluid through the trigger valve  330  may enable control over the position of the piston  320  in the bore  324  by controlling the volume of hydraulic fluid in the bore underneath the piston. 
     It may be desirable to use a trigger valve  330  that is a high speed device; i.e. a device that is capable of being opened and closed at least once per engine cycle. A two-position/two-port valve may provide the level of high speed required. The trigger valve  330  may, for example, be similar to the trigger valves disclosed in the Sturman U.S. Pat. No. 5,460,329 (issued Oct. 24, 1995), for a High Speed Fuel Injector; and/or the Gibson U.S. Pat. No. 5,479,901 (issued Jan. 2, 1996) for a Electro-Hydraulic Spool Control Valve Assembly Adapted For A Fuel Injector. Preferably, the trigger valve  330  may include a solenoid actuator similar to the one shown in  FIG. 7 . The trigger valve  330  may include a passage connecting first passage  326  and second passage  346 , a solenoid, and a passage blocking member responsive to the solenoid. The amount of hydraulic fluid in the bore  324  may be controlled by selectively blocking and unblocking the passage in the trigger valve  330 . Unblocking the passage through the trigger valve  330  enables hydraulic fluid in the bore  324  and the first passage  326  to be transferred to the accumulator  340 . 
     An electronic valve controller  500  may be used to control the position of the moveable portion of the trigger valve  330 . By controlling the time at which the passage through the trigger valve is open, the controller  500  may control the amount of hydraulic fluid in the bore  324 , and thus control the position of the piston  320 . 
     With regard to a method embodiment of the invention, the system  300  may operate as follows to control valve actuation. The system  300  may be initially charged with oil, or some other hydraulic fluid, through an optional check valve  350 . Trigger valve  330  may be kept open at this time to allow oil to fill passages  348 ,  346 , and  326 , and to fill bore  324 . Once the system is charged, the controller  500  may close the trigger valve  330 , thereby locking the piston  320  into a relatively fixed position based on the volume of oil in the bore  324 . Thereafter, the controller  500  may determine a desired level of valve actuation and determine the required position of the piston  320  to achieve this level of valve actuation. The controller  500  may then selectively open the trigger valve  330  so that oil is free to escape from the bore  324  as the motion means  100  forces the piston  320  into the bore. If the motion means is not in position to force the piston  320  downward, opening the trigger valve  330  may result in the addition of hydraulic fluid to the bore  324 . Once the trigger valve  330  is closed again, the piston  320  is locked and the motion means  100  may then apply a fixed displacement motion to the pivoting bridge  200 , while the pivoting bridge is supported on one end by the piston  320 . The cycle of opening and closing the trigger valve may be repeated once per engine cycle to selectively lose a portion or all of a valve event. 
     The system  300  may be designed to provide limp home capability should the system develop a hydraulic fluid leak. Limp home capability may be provided by having a piston  320 , piston spring  322 , and bore  324  of a particular design. The combined design of these elements may be such that they provide a piston position which will still permit some level of valve actuation when the bore  324  is completely devoid of hydraulic fluid. The system  300  may provide limited lost motion, and thus limp home capability, in three ways. Limiting the travel of the piston  320  in its bore  324  may limit lost motion; limiting the travel of the accumulator piston  341  in the accumulator bore  344  may limit lost motion; and contact between the pivoting bridge surface  220  and the housing  310  may limit lost motion Limiting lost motion through contact between the pivoting bridge surface  220  and the housing  310  may be facilitated by making surface  220  wider than the bore  324  so that the outer edges of the surface  220  may engage the housing  310 . 
     Alternative designs for the pivoting bridge  200  which fall within the scope of the invention, are shown in  FIGS. 2, 3 and 5 . The pivoting bridge  200  shown in  FIG. 3  is a Y-shaped yoke that includes two surfaces  230  for contacting two different valve stems (not shown). The pivoting bridge  200  shown in  FIG. 5  includes a cam roller  210  for direct contact with a cam. 
     In alternative embodiments of the invention, the trigger valve  330  need not be a solenoid activated trigger, but could instead be hydraulically or mechanically activated. No matter how it is implemented, the trigger valve  330  preferably may be capable of providing one or more opening and closing movements per cycle of the engine and/or one or more opening and closing movements during an individual valve event. 
     An alternative embodiment of the system  300  of  FIG. 1  is shown in  FIG. 4 , in which like reference numerals refer to like elements. With reference to  FIG. 4 , the piston  320  may be slidably provided in a bore  324 , and biased upward by a piston spring  322 . The bore  324  may be charged with hydraulic fluid provided through a fill passage  354  from a fluid source  360   b . Hydraulic fluid may be prevented from flowing back out of the bore  324  into the fill passage  354  by a check valve  352 . 
     Hydraulic fluid in the bore  324  may be selectively released back to the fluid source  360   b  through a trigger valve  330 . The trigger valve  330  may communicate with the bore  324  via a first passage  326 . The trigger valve  330  may include a trigger housing  332 , a trigger plunger  334 , a solenoid  336 , and a plunger return spring  338 . Selective actuation of the solenoid  336  may result in opening and closing the plunger  334 . When the plunger  334  is open, hydraulic fluid may escape from the bore  324  and flow back through the trigger valve and passage  346  to the fluid source  360   b . The selective release of fluid from the bore  324  may result in selective lowering of the position of the piston  320 . When the plunger  334  is closed, the volume of hydraulic fluid in the bore  324  is locked, which may result in maintenance of the position of the piston  320 , even as pressure is applied to the piston from above. 
     With reference to  FIG. 6 , in which like reference numerals refer to like elements, a preferred variable valve actuation system  10  embodiment of the invention is shown. In  FIG. 6 , the means for providing valve actuation motion  100  is shown as a cam. As with the previously described embodiments, the motion means  100  may include various valve train elements, such as a cam (shown in  FIG. 6 ), or a rocker arm or lever pushrod (shown in  FIG. 1 ). A fixed valve actuation motion may be provided by the motion means  100  via one or more lobes  112  on the cam. 
     Motion means  100  may contact a pivoting lever (bridge)  200  at a centrally defined point  211 . A cam roller  210  may be provided at the central point. The lever  200  may also include a pinned end  220  connected to an adjustable piston  320 , and a contact stem  205  with a surface  230  in contact with a valve stem  400 . Depending upon the needs of the valve actuation system, the lever  200  may be Y-shaped so that a single lever is used to actuate two engine valves. Furthermore, bridges (not shown in  FIG. 6 ) may be used at either the valve contact end  230  or the pinned end  220  of the lever  200 , so that two or more engine valves are linked to one piston  320 . 
     Valve springs  410  may bias the valve stem  400  upward and cause the adjustable piston  320  to be slidably biased downward into a bore  324  provided in the housing  310 . As in the embodiment shown in  FIG. 1 , the housing  310  may further support a trigger valve  330 , an accumulator  340 , and a piston spring  322 . References throughout the specification to the housing  310  should be interpreted to cover any means of containing the system  10 , whether the containing means is a separate housing or a preexisting engine component such as an engine head or valve cover. 
     In addition to the foregoing elements, which are also included in the embodiment of the invention shown in  FIG. 1 , the embodiment shown in  FIG. 6  may also include an electronic valve controller  500  including specialized control algorithms, an initial start mechanism  600 , an optional modified low pressure (i.e. less than a couple hundred psi) hydraulic supply system  700 , and a Self Adjusting Valve Catch (SAVC)  800 . Detailed discussion of these additional elements is provided below. 
     The housing  310  may include multiple passages therein for the transfer of hydraulic fluid through the system. A first passage  326  in the housing  310  may connect the bore  324  with the trigger valve  330 . A second passage  346  may connect the trigger valve  330  with the accumulator  340 . A third passage  348  may connect the accumulator  340  with an hydraulic fluid supply system  700  through a check valve  350 . In an alternative embodiment of the invention, the check valve  350  may not be required. 
     The piston  320  may be connected by a pin  360 , or other connection means to the lever  200 , which is biased upward by the spring  322 . The biasing force provided by the spring  322  may be sufficient to hold the lever  200  against the motion means  100 , but not so large as to cause engine valve float. The spring  322  may comprise a single spring directly under the lever  200  or two or more springs laterally spaced from the longitudinal axis of the lever. 
     The accumulator  340  may include an accumulator piston  341  slidably disposed in an accumulator bore  344  and biased downward by an accumulator spring  342 . Low pressure hydraulic fluid (in the preferred embodiment) that passes through the trigger valve  330  may be stored in the accumulator  340  until it is used to refill the bore  324 . 
     Linear displacement may be provided by the motion means  100  to the lever  200 . Displacement provided to the lever  200  may be transmitted through surface  230  of the contact stem  205  to the valve stem  400 . With reference to  FIG. 8 , the surface  230  of the contact stem  205  may have a dual radius of curvature so as to assist in self-correction of engine valve displacement differences that result from machining and assembly tolerances. The contact stems  205  may also serve to decelerate the lever  200  during Early Valve Closing or Centered Lift operational modes by contacting the SAVC  800  just prior to seating of the engine valve. 
       FIG. 9 , in which like reference numerals refer to like elements, is a detailed pictorial illustration of a preferred embodiment of a Y-shaped lever  200  that may be used with the system shown in  FIG. 6 . The lever  200  shown in  FIG. 9  includes laterally extending flanges  250  which are adapted to receive laterally spaced springs (shown in  FIG. 6 ). The Y-shaped lever  200  may include a relatively wide space to accommodate a cam roller (not shown) and a recess  212  to accommodate pinning the piston (not shown) to the pinned end  230  of the lever. 
     With renewed reference to  FIG. 6 , the valve actuation motion that is transmitted by the motion means  100  to the valve stem  400  via the lever  200  may be controlled by controlling the position of the pinned end  220  of the lever. Given the input of a fixed downward motion by the motion means  100 , if the position of the pinned end  220  of the lever is lowered, then the downward motion experienced by the valve stem  400  is decreased relative to what it would have been otherwise. Thus, by selectively lowering the position of the pinned end  220  through adjustment of the piston  320 , motion imparted by the motion means  100  to the lever  200  may be selectively “lost.” 
     With continued reference to  FIG. 6 , as with the system shown in  FIG. 1 , the displacement experienced by the valve stem  400  may be controlled by controlling the release of the fluid in the bore  324  that holds the piston  320  in place at a selective time during a downward displacement imparted by the motion means  100 . During such a downward displacement, the piston  320  pressurizes the hydraulic fluid in bore  324  beneath the piston. The (now high pressure) hydraulic fluid extends from the bore  324  through the first passage  326  to the trigger valve  330 . Thus, selectively timed opening of the trigger valve  330  causes the piston  320  to slide into the bore  324  and results in the loss of the motion imparted by the motion means  100 . 
     A normally open (or closed) high-speed solenoid trigger valve  330  permits lost motion at the pinned end  220  of the lever  200  or prevents the loss of motion transmitted to the engine valve(s)  400  if it is activated by current from the engine controller  500  (which may contain a microprocessor linked to the engine fuel injection ECM). It may be desirable to use a trigger valve  330  that is a high speed device; i.e. a device that is capable of being opened and closed at least once during an engine cycle, and even as rapidly as on a cam lobe-by-lobe basis. Such rapid trigger valve actuation permits high speed valve actuation, such as is required for two cycle compression release engine braking (where a compression release event occurs each time the engine piston rotates through top dead center position). The trigger valve  330  may, for example, be similar to the trigger valves disclosed in the Sturman U.S. Pat. No. 5,460,329 (issued Oct. 24, 1995), for a High Speed Fuel Injector; and/or the Gibson U.S. Pat. No. 5,479,901 (issued Jan. 2, 1996) for a Electro-Hydraulic Spool Control Valve Assembly Adapted For A Fuel Injector. The trigger valve  330  may include a passage connecting the first passage  326  and the second passage  346 , a solenoid, and a passage blocking member responsive to the solenoid. The amount of hydraulic fluid in the bore  324  may be controlled by selectively blocking and unblocking the passage in the trigger valve  330 . Unblocking the passage through the trigger valve  330  enables hydraulic fluid in the bore  324  and the first passage  326  to be transferred to the accumulator  340 . 
     The preferred trigger valve  330  that may be used with the invention is shown in  FIG. 7 . The trigger valve  330  may include an upper solenoid actuator  336  and a lower piston  334 . A central pin  331  provided in the upper solenoid actuator  336  may be biased downward by an upper spring  333  into contact with the lower piston  334 . The lower piston  334  may be biased upward by a lower spring  335  into contact with the central pin  331 . When the trigger valve  330  is deactivated, the bias of the lower spring  335  overcomes the bias of the upper spring  333 , and the lower piston  334  opens to allow the flow of hydraulic fluid from the first passage  326  to the second passage  346 . When the trigger valve  330  is activated, the central pin  331  and the armature  329  are magnetically attracted downward, allowing the lower piston  334  to be displaced downward onto its seat  339 , and thereby preventing hydraulic communication between the first and second passages  326  and  346 . 
     With renewed reference to  FIG. 6 , the system  10  may operate as follows to control valve actuation. The system may be initially charged with oil, or some other hydraulic fluid, through a check valve  350  (this check valve may be eliminated in an alternative embodiment). The trigger valve  330  may be kept open at this time to allow oil to fill the first passage  326  and the piston bore  324 . Once the system is charged, the controller  500  may close the trigger valve  330 , thereby locking the piston  320  into a relatively fixed position based on the volume of oil in the bore  324 . Thereafter, the controller  500  may determine a desired level of valve actuation and determine the required position of the piston  320  to achieve this level of valve actuation. 
     During the time that the motion means  100  is applying a force to the lever  200 , the controller  500  may open the trigger valve  330  at a selective time, which results in the piston  320  being forced down into the bore  324 , which in turn drives fluid from the bore. Hydraulic fluid (oil) that is driven from the bore  324  as a result of lost motion operation may pass through the trigger valve  330  to the low pressure accumulator gallery that includes one or more individual accumulators  340  fed with cylinder head port oil. The accumulator gallery is connected to one or more accumulators  340  in order to conserve displaced fluid and promote refilling of the bore  324  upon the next cycle of engine valve actuation Bleed orifices or diametrical clearances may be provided in the low pressure section of the accumulator  340  and the valve catch  800  to provide cooling of the system through gradual cycling of the fluid in the system. 
     After the piston  320  completes the loss of the motion imparted by the motion means  100  fluid pressure from the accumulator  340  may force the piston  320  back upward as the motion means returns to its base state (i.e. base circle for a cam). 
     With continued reference to  FIG. 6 , the system  10  may also be designed to provide limp home capability should an hydraulic fluid leak occur. Limp home capability may be provided by having a piston  320  and bore  324  of a particular design, an accumulator piston and accumulator bore of a particular design, or a lever  200  and a housing  310  of a particular design. The combined design of these elements may be such that they provide a piston position which will still permit some level of main event valve actuation and possibly a lower level of valve actuation for some auxiliary event(s) when the bore  324  loses hydraulic fluid pressure. Limp home capability may also be provided by an external fixed stop used when the system  10  contains insufficient hydraulic fluid. 
       FIG. 6A  shows an alternative embodiment of the invention that is very similar to that shown in  FIG. 6 . In  FIG. 6A , a passage connecting the first passage  326  and the second passage  346  is added. A check valve  350  is provided in this additional passage so that fluid flow may only occur from the second passage  346  to the first passage  326 . This additional passage may be used to provide a constant feed of hydraulic fluid to the piston bore  324  regardless of the operational state of the trigger valve  330 . 
     Reference will now be made in detail to the self adjusting valve catch (SAVC) portions of the present invention. The following described valve catch may be used in the various embodiments of the invention, such as those shown in  FIGS. 6 and 11-19 , in the position of valve catch  800 . 
       FIG. 10  is a cross-section of the valve catch portion of the present invention. The valve catch  800  includes an upper member  810  and a lower member  820 . The upper member  810  may include an upper piston  812  and an upper piston spring  814  which biases the upper piston downward. The lower member  820  may include a sleeve  822 , a cup piston  824 , a central pin  826 , a lower spring  828 , a throttling disk  830 , a cap  836 , and a retaining member  838 . The throttling disk  830  may include a center passage  832  and an off-center passage  834 . The cup piston  824  may include a lower surface  825  adapted to contact a contact pin, another feature of the rocker lever, or a valve stem directly. It should be noted that in an alternative embodiment the upper member  810  and the lower member  820  may be fixedly connected together. 
     The components in  FIG. 10  are in the position they would assume when the engine valve  400  is seated, i.e. between valve events. The upper piston spring  814  has pushed the upper piston  812  down into contact with the lower member  820  and has pushed both the upper and lower members down until the cup piston  824  has contacted the Y-bridge  200  or engine valve  400  as appropriate. Hydraulic fluid leaks past the outer diameter of the upper piston  812  to fill the area around the upper piston spring  814 . The upper piston  812  is hydraulically locked and cannot move quickly. When the engine valve  400  opens, low pressure fluid in the supply passage  835  will cause the lower member  820  to move downward until the sleeve  822  contacts the retaining member  838 . Fluid will also flow in through the center of the cap  836 , past the throttling disk  830  and push the cup piston  824  down until it hits the end of the sleeve  822 . Leakage past the upper piston  812  is so slow that the upper piston will have virtually no movement during the time the engine valve  400  is off of its seat. When the engine valve  400  is closing and approaches its seat, the valve stem or lever  200  will first hit the cup piston  824 , pushing the lower member  820  upward until the cap  836  hits the upper piston  812 . Continued engine valve motion will force the cup piston  824  upward within the sleeve  822 , forcing fluid out of the holes in the throttling disk  830  and back into the supply passage  835 . The restricted flow through the holes in the throttling disk  830  will produce an internal pressure in the lower member  820 , slowing the engine valve motion. As the engine valve gets closer to its seat, the central pin  826  will start to block the central orifice  832 , further restricting fluid flow there through and controlling the seating velocity. The stroke of the cup piston  824  within the lower member  820  and the diameter of orifices  832  and  834  can be adjusted to produce the desired seating velocity with a large variation in valve closing velocities. 
       FIGS. 11 and 12  are top plan views of various combinations of lever arms  200  that may used in accordance with various embodiments of the invention.  FIG. 11  shows a Y-shaped intake lever  200   a  and a Y-shaped exhaust lever  200   b  disposed over intake and exhaust valves  400 .  FIG. 12  shows two individually actuated intake levers  200   a  and a Y-shaped exhaust lever  200   b . The individually actuated intake levers  200   a  permit the introduction and control of intake swirl into the cylinder by slightly advancing or delaying the opening or closing of one of the intake levers. 
     An alternative embodiment of the invention is shown in  FIGS. 13 and 14 , in which like reference numerals refer to like elements. With reference to  FIGS. 13 and 14 , a bridge  420  is disposed between the lever  200  and two valve stems  400 . The bridge  420  permits the valve actuation provided by a single bar-shaped lever  200  to be transmitted to two engine valves  400 . 
     Another alternative embodiment of the invention is shown in  FIGS. 15 and 16 , in which like reference numerals refer to like elements. With reference to  FIGS. 15 and 16 , a rear bridge  240  is connected to a piston  320  by a pin  360 . The bridge  240  permits a single piston  320  to be used to adjust the vertical position of the pinned end of two levers  200 . 
     Still another alternative embodiment of the invention is shown in  FIGS. 17 and 18 , in which like reference numerals refer to like elements. With reference to  FIGS. 17 and 18 , the location of the cam roller  210  has been moved to the end of the lever  200 , and the piston  320  is pinned to the lever at a point between the cam roller and the contact stem  205 . Furthermore, the piston  320  resides in an overhead assembly. 
     The lower control piston  320 ′ shown in  FIG. 17  may be used instead of the control piston  320  in an alternative embodiment of the invention. The lower control piston  320 ′ may be located on the same side of the lever as the cam  100  if the position of the lower control piston  320 ′ is dictated by fluid flow to and from a chamber located above the control piston as opposed to below the control piston. 
     Still another alternative embodiment of the invention is shown in  FIG. 19 , in which like reference numerals refer to like elements. The piston  320  and the lever  200  may be connected using a ball and socket arrangement. Although the ball is shown as part of the piston  320  and the socket is shown as part of the lever  200 , it is appreciated that the ball could be integrally formed with the lever and the socket could be formed in the piston. 
     The Initial Start Mechanism and Hydraulic Fluid Supply System 
     The VVA systems shown in  FIGS. 6-19  each need to be charged with hydraulic fluid in order to operate properly. It is typically the case, however, that the hydraulic fluid contained in these systems will largely drain out once the engine is shut off. The recharging of the system with hydraulic fluid upon initial start of the engine may take some time, during which there will be no “hydraulically actuated” valve motion. Thus, there is a need for a system that accelerates the process of charging the VVA systems with hydraulic fluid, and/or for a system that provides some fixed level of valve actuation even when the VVA systems are devoid of hydraulic fluid. Applicants have developed several initial start mechanisms  600  and several modified hydraulic fluid supply systems  700  in an attempt to meet the foregoing needs. 
     Two general types of initial start mechanisms (ISMs)  600  are disclosed herein. The first type of ISMs provide a fixed stop near the pinned end  220  of the lever  200 . In these systems, the fixed stop may be automatically removed once the overall VVA system is charged with hydraulic fluid. These types of ISMs are depicted in  FIGS. 20-26 . The second type of ISMs are those that lock the piston  320  into a fixed position until the overall VVA system is charged with hydraulic fluid. These ISMs are depicted in  FIGS. 27-41 and 81-83 . 
     With reference to  FIG. 20 , an ISM  600  is installed below the pinned end  220  of the lever  200 . The ISM  600  includes an ISM piston  610  slidably disposed in a bore  612  that receives oil from the low pressure supply  700  (i.e. the engine) used to charge the VVA system. The bore  612  is vented to atmosphere by passage  640 . The ISM piston  610  is biased by a spring  614  such that the piston body  616  is directly below the locking shaft  620  when there VVA system is devoid of hydraulic fluid. When the ISM piston  610  is in this position it provides a bottom support for the locking shaft  620 , thereby permitting the locking shaft to support the pinned end  220  of the lever  200  when the piston  320  is incapable of doing so. 
     The locking shaft  620  is biased upward into contact with the lever  200  by the piston spring  322 . When the locking shaft  620  is supported by the piston body  616  it provides a fixed stop for the lever  200 . The length of the locking shaft may be selected such that with the exception of the main intake and main exhaust events, the motion of all cam lobes is lost. Such actuation is typically preferred during engine starting. When the piston body  616  is not below the locking shaft  620 , however, the locking shaft is free to be displaced downward against the bias of the piston spring  322  into the bore  612 . 
     After initial starting of the engine, hydraulic fluid is supplied to the bore  612 . This hydraulic fluid acts on the ISM piston plunger head  618  and forces the ISM piston  610  back into the bore  612  against the bias of the spring  614 . Movement of the ISM piston  610  is possible due to the venting of hydraulic fluid past the piston through the passage  640 . As the ISM piston  610  slides back, the bottom support for the locking shaft  620  is removed, thereby eliminating the locking shaft&#39;s ability to act as a fixed stop. The continued flow of hydraulic fluid into the VVA system passes through the trigger valve  330  and into the piston bore  324 . At this point the trigger valve  330  may be closed, and support for the lever  200  may be provided by the piston  320 . 
     With continued reference to  FIG. 20 , the ISM  600  may also be provided with an optional valve  630 . The optional valve  630  may provide a limp-home mode of operation for the VVA system when there is some hydraulic pressure, but not sufficient pressure for the system to operate properly. When the valve  630  is closed, low pressure hydraulic fluid may leak past the plunger head  618  and the piston body  616  into the rear portion of the bore  612 . This leakage may cause a buildup of hydraulic pressure behind the ISM piston  610  causing it to move forward in the bore  612  until it provides a support for the locking shaft  620 . 
     A similar system to that shown in  FIG. 20  is shown in  FIG. 21 , in which like reference numerals refer to like elements. With reference to  FIG. 21 , the ISM piston  610  is slidably disposed in the bore  612  such that it provides a fixed support for the piston  320  when the VVA system is devoid of hydraulic fluid. Application of hydraulic fluid to the system through the trigger valve  330  and into the bore  612  not only charges the system with fluid, but also pushes the ISM piston  610  back into the bore  612  so that the piston  320  is free to slide to the bottom of the bore  324 . 
     With reference to  FIG. 22 , the ISM  600  is capable of providing a fixed stop for a plurality of levers  200 . The ISM  600  includes sliding bars  670  that are biased by the bar springs  672  into a position that the raised portions  673  are directly underneath the levers  200 . When in this position, the sliding bars  670  provide fixed stops for the levers  200  such that the main exhaust and main intake valve events are transmitted from the cams to the engine valves even when the VVA system is devoid of hydraulic fluid. 
     Application of hydraulic fluid to the VVA system results in the flow of fluid into the bore  678 . The hydraulic fluid in the bore  678  pushes the inclined piston  674  upward against the bias of the spring  676  and into contact with the sliding bars  670 . The inclined end faces of the sliding bars  670  and the inclined face of the piston  674  slide against one another, causing the sliding bars to be laterally displaced toward the bar springs  672 . As the sliding bars  670  are displaced, the levers  200  ride down from the raised portions  673  on the bars until the levers are free to pivot on the pistons  320  (not shown). 
     With continued reference to  FIG. 22 , the sliding bars  670  may be aligned using a guide rail or grooves  675  running the length of the cylinder head. The guide rail or grooves  675  may mate with an inverse feature provided along the bottom surface of the sliding bars  670 . 
     With reference to  FIG. 24 , the sliding bars may be provided with a small amount of clearance  679  beneath the raised portions  673 . The clearance  679  may permit deflection x of the sliding bar as the lever  200  is pressed down on the bar during a valve event. It is anticipated that the desired deflection x of the bar  670  is on the order of a few hundredths of a millimeter. Such deflection may provide a cushioning effect as the lever  200  impacts the bar  670  during a valve event. 
     With reference to  FIG. 23 , an alternative embodiment of the ISM  600  is shown. The operation of the ISM  600  shown in  FIG. 23  is the same as that shown in  FIG. 22 , with the exception of the use of two sliding bars  670  and a centrally located inclined piston  674 . 
     With reference to the embodiments shown in both  FIGS. 22 and 24 , it is anticipated that the height of the fixed stop required for an intake valve arrangement and that for an exhaust valve arrangement will be different. The same sliding bar  670  may be used for both intake and exhaust valve arrangements, however, provided that the height of the surfaces on which the bars slide are different. An intake lever could be positioned over a slot having a lesser depth for receipt of a first sliding bar  670 . An exhaust lever could be positioned over a slot having a greater depth for receipt of a second sliding bar  670 . The same size sliding bar  670  may be used for both the intake and the exhaust levers because the individualized depth of the slots in which the bars ride controls the height of the fixed stop provided by the sliding bars. This feature eliminates the possibility that the wrong sliding bar will be used with the intake or exhaust valve arrangement. 
     With reference to  FIG. 25 , in which like reference numerals refer to like elements shown in other figures, a fixed stop is provided for the lever  200  in the form of a hinged toggle  650 . The toggle  650  is pivotally mounted and biased into an upright position by the toggle spring  654 . An upright shaft  660  is biased upward into the toggle  650  by fluid pressure underneath the shaft. The toggle  650  and the upright shaft  660  may have mating inclined faces that are adapted to slide against each other. 
     In its upright position, the toggle  650  abuts a boss  202  extending from the lever  200 . In this position the toggle  650  provides a support for the pinned end  220  of the lever  200 . It is appreciated that a second boss could extend from the other side lever  200  and the toggle could be design to engage the bosses on both sides of the lever when the toggle is in an upright position. 
     The toggle  650  may be pivoted out of its upright position when the VVA system is charged with hydraulic fluid. Application of hydraulic fluid to the system results in the flow of fluid into the bore  612 . The hydraulic fluid in the bore  612  may force the upright shaft  660  upwards so that the inclined faces of the toggle  650  and the shaft meet. As the shaft continues to move upward, it causes the toggle  650  to pivot counter-clockwise against the bias of the toggle spring  654 . Eventually the toggle  650  is sufficiently pivoted that it no longer provides a support for the boss  202 , at which point the vertical position of the pinned end  220  of the lever  200  is determined by the position of the piston  320 . 
     With reference to  FIGS. 27 and 28 , another embodiment of an ISM  600  that is adapted to lock the piston  320  into a fixed position is disclosed. The ISM  600  includes an upright piston  690  (which may be the system accumulator elsewhere labeled as  340 ) disposed in an upright bore  695 , piston bias springs  691  and  692 , sliding member  693 , and sliding member bias spring  694 . 
     When the engine is off, hydraulic fluid may drain from the upright bore  695 , permitting the bias springs  691  and  692  to push the upright piston  690  downward into its seat. Positioning of the upright piston  690  in its seat forces the sliding member  693  to move against the bias of the spring  694  such that the raised portion  696  of the sliding member is underneath a boss  321  provided on the piston  320  (or alternatively on the lever  200 ). While in this position, the sliding member  693  provides a fixed stop for the piston  320  to ride against. The height of the fixed stop provided by the sliding member  693  may be preselected to provide some level of valve actuation when the VVA system is devoid of hydraulic fluid. 
     As the engine is started, hydraulic fluid flows into the upright bore  695 , which in turn forces the upright piston  690  to move upward against the bias springs  691  and  692 . As the upright piston  690  moves upward, the sliding member  693  is permitted to slide towards the upright piston under the influence of the bias spring  694 . The ISM  600  is designed such that once the upright piston attains its uppermost position, the raised portion  696  of the sliding member  693  will no longer be underneath the boss  321 . This permits the piston  320  to be raised and lowered freely for VVA actuation upon the charging of the system with hydraulic fluid. 
     Another embodiment of the ISM portion of the present invention is shown in  FIG. 29 . With reference to  FIG. 29 , a control piston  320  is shown with a castellated collar disposed around it. Mating castellations may be provided on the piston  320  and the collar  323 . When the collar  323  is positioned such the castellations thereon mate with those of the piston  320 , the piston is provided with a full range of vertical movement. Alternatively, if rotated by a rotation means  325 , the collar  323  may provide a fixed stop for the piston  320  (to be used during initial starting or limp-home operation). 
     The embodiment of the ISM portion of the present invention that is shown in  FIG. 30  is similar to that shown in  FIG. 25 . With reference to  FIG. 30 , a fixed stop is provided for the control piston  320  in the form of a hinged toggle  650  that may support a piston boss  321 . The toggle  650  is pivotally mounted on a toggle base  652  and weighted (or spring biased) to rotate clockwise when the end  651  is not held down by the upright shaft  660 . 
     When the VVA system is devoid of hydraulic fluid, the upright shaft  660  (which may be provided by an upper extension of the accumulator  340 ) is in the position shown by the phantom lines in  FIG. 30 . As the system is provided with hydraulic fluid, the upright shaft  660  is pushed upwards, permitting the toggle  650  to rotate clockwise and freeing the piston  320  to operate with its full range of motion. 
     Yet another embodiment of the ISM portion of the present invention is shown in  FIG. 31 . With reference to  FIG. 31 , a fixed stop is provided for the control piston  320  in the form of a toggle  650  that may support a piston boss  321 . The toggle  650  is designed, weighted and/or spring biased to move out of position from underneath the piston boss  321  when the end  651  is not held down by the upright shaft  660 . In an alternative embodiment, the boss  321  may be provided on the rocker lever  200  instead of the piston  320 . 
     When the VVA system is devoid of hydraulic fluid, the end  651  is held down in the position shown by the upright shaft  660  (which may be provided by an upper extension of the accumulator  340 ). As the system is provided with hydraulic fluid, the upright shaft  660  is pushed upwards, permitting the end  651  to rise and rotate the toggle  650  out of position from underneath the piston boss  321  so that the piston  320  can operate with its full range of motion. 
       FIG. 26  shows an embodiment of the ISM portion of the present invention similar to that shown in  FIG. 31 . With reference to  FIG. 26 , the toggle  650  is biased into the “on” position (shown) by the flat spring  654 . In the on position, the toggle  650  limits the motion of the control piston  320  when the end of the lever  200  contacts the toggle. In an alternative embodiment, this could also be accomplished by a projection on the control piston  320  contacting the toggle  650 . When the system  10  hydraulic pressure increases, the piston  660  (which may be provided by the accumulator piston  341 ) moves upward, overcoming the bias of the flat spring  654  and tipping the toggle  650  out of engagement with the lever  200 . When the system pressure drops, the piston return spring  658  forces the piston  660  back down into its bore, allowing the flat spring  654  to move the toggle  650  back into the engaged position. 
     Should the engine stop with the lever  200  in a depressed position, the flat spring  654  will press the toggle  650  into the side of the lever. As soon as the lever  200  moves as the result of cranking the engine, the toggle  650  will snap into the engaged position. Should the lever  200  move back down before the toggle  650  reaches its most upright position, the toggle will be pushed back down without damage, and will be able to reset the next time the lever rises. 
     With reference to  FIG. 32 , a second general type of ISM  600  is shown. The ISM  600  shown in  FIG. 32  operates by locking the control piston  320  into a fixed position until such time as the overall VVA system is charged with hydraulic fluid. The ISM  600  includes an inner locking piston  680  slidably disposed inside of a control piston  320  and biased downward by a spring  681 . The control piston  320  is slidably disposed in a control piston bore  324  defined by a sleeve  685 . Locking balls  686  are moveable in a space defined by a through-hole in the wall of the control piston  320 , a sleeve recess  687 , and a locking piston recess  688 . 
     When the piston bore  324  is devoid of hydraulic fluid (as it is during start up) the spring  681  extends and forces the inner locking piston  680  to slide downward relative to the control piston  320 . The downward movement of the locking piston  680  forces the locking balls  686  outward into the space defined by the sleeve recess  687  and tie through-hole in the wall of the control piston  320 . This positioning of the locking balls  686  mechanically locks the control piston  320  in a fixed position relative to the sleeve  685 . Thus, when there is no hydraulic fluid in the piston bore  324 , the piston  320  may be automatically locked into a fixed position. 
     As hydraulic fluid flows into the piston bore  324 , the inner locking piston  680  is forced upwards into the control piston  320 . A bleed passage  689  may be provided in the control piston  320  to avoid hydraulic lock of the inner locking piston  680  in the control piston. As the inner locking piston  680  moves upward, it comes to rest against a shoulder provided in the control piston  320 . Any further upward movement of the locking piston  680  causes the control piston  320  to move upward as well. As the control piston  320  moves upward, the curved wall of the control piston recess  687  urges the locking balls  686  into the space defined by the control piston through-hole and the locking piston recess  688 . In this manner, the control piston  320  is unlocked from the sleeve  685  and the piston  320  is free to slide vertically in the piston bore  324 , and it should be noted that the unlocking action of the recess  687  can achieve the same function of unlocking when the control piston  320  and the inner piston  680  move as one unit in the downward direction. 
     With reference to  FIG. 33 , an alternative embodiment of the locking mechanism for the control piston  320  is shown. Like that shown in  FIG. 32 , the ISM  600  shown in  FIG. 33  operates by locking the control piston  320  into a fixed position until such time as the overall VVA system is charged with hydraulic fluid. The ISM  600  includes an inner piston  680  slidably disposed inside of a control piston  320  and biased downward by a spring  681 . The control piston  320  is slidably disposed in a piston bore  324  defined by a sleeve  685 . A locking ring or balls  686  are laterally moveable in the bore  324 . The control piston  320  may include lower walls that are predisposed to deflect inward, but which may be deflected outward by a downward movement of the inner piston  680 . 
     When the piston bore  324  is devoid of hydraulic fluid (as it is during start up) the spring  681  extends and forces the inner piston  680  to slide downward relative to the control piston  320 . The downward movement of the inner piston  680  forces the locking ring or balls  686  outward into the sleeve recess  687 . This positioning of the locking ring  686  mechanically locks the control piston  320 ) in a fixed position relative to the sleeve  685 . Thus, when there is no hydraulic fluid in the piston bore  324 , the piston  320  may be automatically locked into a fixed position. 
     As hydraulic fluid flows into the piston bore  324 , the inner locking piston  680  is forced upwards into the control piston  320 . A bleed passage  689  may be provided in the control piston  320  to avoid hydraulic lock of the inner locking piston  680  in the control piston. As the inner locking piston  680  moves upward, the lower walls of the control piston  320  are once again free to deflect inward. The inward deflection of the control piston walls permits the locking ring  686  to contract and unlock the control piston  320  from the sleeve  685 . 
     Another ISM embodiment of the invention that may be used to lock the control piston  320  into place during initial starting is shown in  FIGS. 34-37 . With reference to  FIGS. 34-37 , the control piston  320  may be provided with one or more side wall recesses  627 . The recesses  627  may be defined by each set of neighboring protrusions  628 . A splined locking ring  621  may surround the control piston  320 . The ring  621  may include a number of splines  622  that are adapted to slide through the recesses  627  provided on the control piston  320 . The ring  621  may also include an arm  623  extending out from the ring and into selective contact with a deactivation piston  624 . The ring  621  may be biased to rotate either clockwise or counter-clockwise under the influence of a spring  626 . 
     When there is little or no hydraulic fluid in the system, the deactivation piston  624  is recessed into the system housing, leaving the arm  623  and the connected locking ring  621  free to rotate under the influence of the spring  626 . During this time, the locking ring  621  is rotated into a position such that the splines  622  on the ring do not mate with the recesses  627  on the control piston  320 . Accordingly, the control piston  320  is locked into an extended position when there is little or no hydraulic fluid in the system. 
     As the system charges with hydraulic fluid, the deactivation piston  624  is pushed upward and into contact with the arm  623 . The upper ramped portion  625  of the deactivation piston engages the arm  623  and rotates the ring  621  back into the position shown in  FIG. 34 . When the ring  621  is in this position, the splines  622  thereon mate with the recesses  627  on the control piston  320  and the control piston is free to slide up and down to effect variable valve actuation. 
       FIGS. 38-40  show yet another ISM  600  that may be used to lock the control piston  320  into an extended position during initial starting. The ISM  600  includes a control piston  320  with side indents  631 . A deactivation piston  624  is located next to the control piston  320 . The deactivation piston  624  may include a dual ramped upper portion  625 . Twin pincer arms  632  may extend from the deactivation piston  624  to the control piston  320 . A spring  633  may bias the locking ends  634  of the pincer arms  631  to close inward and engage the indents  631  on the control piston. 
     With continued reference to  FIGS. 38-40 , when there is little or no hydraulic fluid in the system, the deactivation piston  624  is recessed into the system housing, allowing the pincer arms  632  to engage the control piston  320  and lock it into an extended position. As the system charges with hydraulic fluid during start up, the deactivation piston  624  is pushed upward and into contact with the ends of the pincer arms  632 . The upper ramped portion  625  of the deactivation piston engages the ends of the pincer arms  632  and forces them inward against the bias of the spring  633 . As a result, the locking ends  634  of the pincer arms  632  move outward and disengage the control piston  320  leaving the control piston free to slide up and down to effect variable valve actuation. 
     With reference to  FIG. 41 , another ISM  600  is shown. This ISM includes a control piston  320  with two radially mounted flaps  635  that can move from a retracted position  636  out to an extended position  637 . When the flaps  635  are in the retracted position  636 , the control piston  320  is free to slide vertically for variable valve actuation. When the flaps  635  are in the extended position  637 , the control piston  320  is locked into an extended position for initial start-up actuation. The position of the flaps  635  may be controlled with a rotating ring  639 . The ring  639  is shown in section behind the flaps  635 . The ring  639  may be provided with a non-uniform inner surface that allows the flaps  635  to be extended when the ring is in a first position and retracted when the ring is in a second position. Rotation of the ring  639  between the first and second positions may be controlled using the principles and apparatus described in connection with  FIGS. 34-37  for the rotation of the locking ring shown therein. 
     With reference to  FIG. 81 , in which like reference characters refer to like elements in the other drawings, another embodiment of an ISM  600  that is adapted to lock the piston  320  into a fixed position is disclosed. The piston  320  is provided with one or more internal passages  990  which provide hydraulic communication between the piston bore  324  and an annular indentation  991  provided in the side wall of the piston  320 . A lock piston  992  may be slidably disposed in a lock piston bore provided in the housing  310 . The lock piston bore may intersect the piston bore  324  at a right angle. The lock piston  992  may be biased towards the piston  320  by a lock piston spring  993 . The lock piston  992  may have an outer end adapted to slide into and engage the annular indentation  991 . 
     During engine operation, hydraulic fluid pressure in the piston bore  324  is sufficient to overcome the bias of the lock piston spring  993  and keep the lock piston  992  from engaging the annular indentation  991 . As a result, the piston  320  may move freely in the piston bore  324  under the control of a trigger valve (not shown). At the conclusion of engine operation, hydraulic fluid pressure in the piston bore  324  may decrease as fluid “leaks down”. The decreased hydraulic fluid pressure in the piston bore  324  may cause the piston  320  to retract in the piston bore. At the same time, the decreasing pressure in the piston bore  324  may cause the lock piston spring  993  to push the lock piston  992  towards the piston  320  and into the annular indentation  991 . By engaging the annular indentation  991 , the lock piston  992  locks the piston  320  into a fixed position. The fixed position of the piston  320  may be selected to provide a predetermined level of engine valve actuation. At engine start up, the fixed position of the piston  320  may enable engine valve actuation when there is otherwise insufficient hydraulic pressure in the piston bore  324  for engine valve actuation. After engine start up, increased hydraulic pressure in the piston bore  324  may push the lock piston  992  back out of the annular indentation  991 , thereby unlocking the piston  320 , and enabling the piston  320  to move freely again to provide variable valve actuation. The ISM system  600  shown in  FIG. 81 , as well as in other figures, may also provide a fixed level of engine valve actuation when hydraulic fluid pressure is lost for any reason, regardless of whether or not the engine is in start up mode. 
       FIG. 82  illustrates an alternative embodiment of the ISM system  600 . The system shown in  FIG. 82  is the same as that shown in  FIG. 81  with the following exceptions. In  FIG. 82 , the piston  320  is provided with a diametrical passage in which one or more lock pistons  992  are slidably disposed. The one or more lock pistons  992  may be biased away from the center of the piston  320  by a lock piston spring  993 . An annular indentation  995  may be provided in the side wall of the piston bore  324 . The annular indentation  995  may communicate with a hydraulic fluid supply (not shown) through a lock piston supply passage  994 . The hydraulic fluid supply may be common for the piston bore  324  and the lock piston supply passage  994 . A lock piston drain passage  996  may extend from the diametrical passage housing the lock pistons  992  to a lower portion of the piston  320  outer wall. When hydraulic fluid pressure is low in the piston bore  324 , it may also be low in the lock piston supply passage  994 . During low hydraulic fluid pressure conditions, such as during engine start up, the lock piston spring  993  pushes the lock pistons  992  into the annular indentation  995 , thereby locking the piston  320  into a fixed position. When hydraulic fluid pressure is higher, the lock pistons  992  may be pushed back into the diametrical passage, thereby allowing the piston  320  to move freely for variable valve actuation. Hydraulic fluid which leaks into the space between the lock pistons  992  may drain through the lock piston drain passage  996 . 
       FIG. 83  illustrates a slight modification of the ISM  600  shown in  FIG. 82 . The ISM  600  in  FIG. 83  utilizes a single lock piston  992  as opposed to the multiple lock pistons shown in  FIG. 82 . In all other respects, the system shown in  FIG. 83  is the same as, and operates similarly to, the system shown in  FIG. 82 . 
     A first embodiment of an hydraulic fluid charging system  700  portion of the present invention is shown in  FIG. 42 . The system  700  includes a inlet check valve  701  that may receive hydraulic fluid (oil) from the main engine supply. Oil passing through the inlet check valve  701  passes through an air vent unit  702  to an hydraulic circuit  703 . The hydraulic circuit  703  may pass close to an engine water cooling jacket  715  to remove heat from the oil in the hydraulic circuit  703 . The hydraulic circuit connects to the VVA gallery  713  through the check valve  704  and the inlet pump  705 . The hydraulic circuit  703  may also connect to a bore housing a solenoid or pressure driven valve  710 . A relief valve  714  permits oil to flow from the VVA gallery  713  to the hydraulic circuit  703  as needed. 
     The inlet pump  705  may be mechanically driven and connected to the VVA gallery  713  by a pump outlet  706 . The VVA gallery  713  may be connected to plural passages  348  associated with each VVA system. The last two outlets of the VVA gallery  713  may lead to a bore housing the valve  710 . The valve  710  may include a first internal passage arrangement  711  and a second internal passage arrangement  712 . The bore housing the solenoid driven valve  71   0  may also include two openings connecting the spool valve  710  to a mechanically driven outlet pump  707 . The outlet pump  707  may include an inlet port  708  and an outlet port  709 . 
     The system  700  may be operated as follows to provide a high oil pumping rate to the VVA gallery  713  during engine start-up and a relatively low oil pumping rate during steady-state engine operation. As an initial matter, the inlet pump  705  may be provided with a pump rate of ten (10) units per revolution and the outlet pump  707  may be provided with a pump rate of nine (9) units per revolution. The volume of a “unit” and the pump differential of the inlet and outlet pumps may be adjusted as needed to meet the needs of a particular VVA system. It is only important for this portion of the invention that the pump rate of the inlet pump  705  be greater than the pump rate of the outlet pump  707 . 
     During engine start-up the valve  710  is positioned in its bore such that the second spool valve passage arrangement  712  connects the hydraulic circuit  703  to the inlet  708  of the outlet pump  707  and the outlet  709  of the outlet pump to the VVA gallery  713 . When the valve  710  is so positioned, the VVA gallery  713  receives nineteen (19) units of oil per revolution from the hydraulic circuit  703 . Ten (10) units of oil are provided by the inlet pump  705  and nine (9) units of oil are provided by the outlet pump  707 . 
     After engine start-up, the valve  710  may be activated (or de-activated depending upon the normal position of the valve) so that the first valve passage arrangement  711  connects the VVA gallery  713  to the inlet of the outlet pump  707  and connects the outlet  709  of the outlet pump to the hydraulic circuit  703 . When in this position, the VVA gallery is provided with only one unit of oil per revolution of the pumps  705  and  707 . 
     The system  700  selectively provides a high pumping rate to quickly pressurize the VVA gallery on start-up and a low pumping rate to maintain VVA gallery pressure during steady-state engine operation without excessive parasitic loss (as a result of a high flow rate through the relief valve  714 ). The system  700  also provides a high circulation rate of oil through the heat exchanging portion of the system to control system temperature, and de-aeration of make-up oil to improve bulk modulus of the oil in the system. 
     A second embodiment of an hydraulic fluid charging system  700  is shown in  FIG. 43 . With reference to  FIG. 43 , the system  700  includes a cam  100  with one or more lobes  112 . The cam  100  contacts a piston  720  which is biased into contact with the cam  100  by a spring  722 . The piston  720  is disposed in a bore  725 . The space between the end of the bore  725  and the end of the piston  720  defines a pumping chamber  723 . The pumping chamber  723  communicates with an hydraulic reservoir  724  via a passage  726  that may be provided with a check valve  727 . The pumping chamber  723  may also communicate with a VVA gallery (not shown) through a passage  728  that may be provided with a check valve  729 . The reservoir  724  may receive low pressure hydraulic fluid from the engine oil sump via a passage  730 . A return bypass passage  731  including a check valve  732  may connect the passage  728  with the reservoir  724 . 
     Upon engine starting, cranking of the engine causes the cam  100  to rotate. The rotation of the cam  100  causes the piston  720  to slide back and forth in the bore  725 . The piston  720  may be dimensioned such that its back stroke permits it to draw hydraulic fluid from the reservoir  724  through the passage  726 . The forward stroke of the piston  720  pumps hydraulic fluid past the check valve  729  and through the passage  728  to the VVA gallery. 
     A piston locking sub-system  740  may be provided to maintain the piston  720  in a non-pumping position after the VVA gallery is charged with hydraulic fluid. The locking sub-system includes a pin  741  slidably disposed in a pin bore  742 . The pin bore  742  may include a proximal wide portion and a distal narrow portion. The pin  741  may include portions that mate with the wide and narrow portions of the pin bore  742 . The pin  741  may be biased by a spring  743  toward a bore plug  746 . The pin  741  may include a shaped head  744  adapted to engage a recess  721  provided in the piston  720  and a shoulder  745  against which hydraulic pressure may act. The pin bore  742  communicates with a passage  74   7  connected to the engine&#39;s main oil line or the VVA gallery (not shown). 
     At the conclusion of engine start-up, the engine&#39;s oil pump forces oil into the locking sub-system  740  via the passage  747 . This oil may be used to refill the reservoir  724  and to activate the locking sub-system  740 . The oil in passage  747  acts on the shoulder  745  driving the pin  741  against the bias of the spring  743  toward the pin  720 . As the pin  741  moves, the shaped head  744  engages the recess  721  in the piston  720 , thereby locking the piston  720  into a position removed from the cam  100 . Upon engine shut-off, oil drains from the passage  747  allowing the pin  741  to disengage the recess  721  and unlock the piston  720 . 
     The pin bore  742  intersects the piston bore  725  such that neither end of the piston  720  is capable of stroking past the pin bore  742 . This may prevent the piston  720  from being trapped in a locked position within the piston bore  725 , or in an extended position against the cam  100 . 
     It is appreciated that in alternative embodiments, the piston locking sub-system  740  may be provided with a pin  741  that is either stepped (as shown) or uniform (not shown). It is also appreciated that the pin  741  could be replaced by an approximately semicircular ring (shown in  FIG. 44 ) residing in an annulus cut into the piston bore  725 . 
     A third embodiment of the hydraulic fluid charging system  700  portion of the present invention is shown in  FIG. 46 . With reference to  FIG. 46 , the system  700  includes an inlet hydraulic fluid port  759 , check valves  762 , an exit check valve  729 , a pumping piston  761 , a piston bias spring  765 , a fluid reservoir  760 , a solenoid controlled valve  763 , an air bleed tube  758 , and a bleed tube check valve  764 . 
     In the system  700  shown in  FIG. 46 , the pumping piston  761  may be driven by a cam (not shown) so that it moves upward and back repeatedly within the bore housing it. The piston bias spring  765  is included to ensure that the piston  761  follows the contour of the cam (not shown) used to drive it. The solenoid controlled valve  763  is placed in a hydraulic bypass circuit bracketing the pumping piston  761 . The solenoid controlled valve  763  is maintained in an open position during normal engine operation to negate parasitics, and a closed position during engine start up. During normal running, the system  700  is filled with hydraulic fluid ready for the next start. 
     With continued reference to  FIG. 46 , after engine shut down the check valves  762  prevent the hydraulic fluid in the reservoir  760  from leaking out. Upon engine start up, the reciprocal motion of the pumping piston  761  is resumed. Because the reservoir  760  is full of hydraulic fluid and in close proximity to the pumping piston  761 , the piston can immediately draw fluid to charge the VVA system  300 . The bleed tube check valve  764  permits equalization of the pressure in the reservoir  760  when fluid is drawn from it on start up. 
     A fourth embodiment of the hydraulic fluid charging system  700  portion of the present invention is shown in  FIG. 47 . With reference to  FIG. 47 , the system  700  includes an inlet hydraulic fluid port  759  from the engine&#39;s oil sump, check valves  762 , an exit check valve  729 , a pumping piston  761 , a piston bias spring  765 , and a fluid reservoir  760 . 
     In the system  700  shown in  FIG. 47 , the pumping piston  761  may be driven by a cam (not shown) so that it moves upward and back repeatedly within the bore housing it. The operation of the system  700  shown in  FIG. 47  is similar to that shown in  FIG. 46 . The reservoir  760  is filled with fluid during normal operation and is maintained full by the check valves  762  when the engine is shut down. Upon engine start up, the displacement of the pumping piston  761  draws hydraulic fluid from the reservoir  760  and pumps it to the VVA system  300 . The system  700  is disabled automatically as a result of selecting a piston bias spring  765  with a particular biasing strength. The bias spring  765  provides enough force to keep the pumping piston  761  in contact with the cam initially. Once the pressure in the hydraulic circuit underneath the pumping piston  761  reaches normal operating levels, however, the bias of the spring  765  is insufficient to force the pumping piston  761  down. Thus, once normal operating pressure is achieved in the VVA system  300 , the pumping piston  761  will be maintained up out of contact with the cam used to drive it. 
     A fifth embodiment of the hydraulic fluid charging system  700  portion of the present invention is shown in  FIG. 48 . With reference to  FIG. 48 , the system  700  includes an inlet hydraulic fluid port  759 , a check valve  762 , a fluid reservoir  760 , a solenoid controlled valve  763 , and a compressed gas bladder  766 . This embodiment uses the combination of the compressed gas bladder  766  and the solenoid controlled valve  763  to selectively force hydraulic fluid in the reservoir  760  into the VVA system  300  upon engine start up. 
     A sixth embodiment of the hydraulic fluid charging system  700  portion of the present invention is shown in  FIG. 49 . With reference to  FIG. 49 , the system  700  includes an inlet hydraulic fluid port  759 , a check valve  762 , a fluid reservoir  760 , a solenoid controlled catch  769 , a diaphragm  766 , piston  767 , and a spring  768 . The spring  768  biases the diaphragm  766  into a position that forces hydraulic fluid out of the reservoir  760  and into the VVA system  300  via the passage  728 . This embodiment uses the combination of the spring biased diaphragm  766  and the solenoid controlled catch  769  to force hydraulic fluid in the reservoir  760  into the VVA system  300  upon engine start up. 
     A seventh embodiment of the hydraulic fluid charging system  700  portion of the present invention is shown in  FIG. 50 . With reference to  FIG. 50 , the system  700  includes an inlet hydraulic fluid port  759 , check valves  762 , an exit check valve  729 , a cylindrical fluid reservoir  760 , an electric motor  772 , a screw shaft  771 , and a piston  770 . In this embodiment, upon engine start up the electric motor  772  drives the screw shaft  771  to force the piston  770  through the reservoir  760  which results in the hydraulic fluid in the reservoir  760  being forced into the VVA system  300  via the passage  728 . 
     An eighth embodiment of the hydraulic fluid charging system  700  portion of the present invention is shown in  FIG. 51 . With reference to  FIG. 51 , the system  700  includes a housing with an inlet hydraulic fluid port  759  connected through a check valve  762  to a fluid reservoir  760 . The fluid reservoir  760  is connected through a second check valve  762  to a pumping cylinder  774  in which a pumping piston  773  is disposed. The pumping piston  773  is biased upward by a first spring  775  into a lever  776 . The lever  776  pivots on a fulcrum  777  in response to the rotation of a cam  110 . The lever  776  is biased into contact with the cam  110  by a second spring  778 . The pumping cylinder  774  is also connected through an exit check valve  729  with an outlet passage  728 . 
     With continued reference to  FIG. 51 , the motion of the cam  110  is used to supply hydraulic fluid to the VVA system  300 . The motion of the cam  110  causes the lever  776  to pivot on the fulcrum  777  and pump the pumping piston  773  up and down in the pumping cylinder  774 . This pumping action draws oil from the reservoir  760  and pumps it into the VVA system  300  via the outlet passage  728 . The fluid charging system  700  recharges using engine oil pressure from the inlet passage  759 . The reservoir  760  retains this charge of fluid as a result of placement of the first check valve  762  located in the inlet passage  759 . During normal engine operation, the combined force of the first spring  775  and the oil pressure in the pumping cylinder  774  are sufficient to overcome the bias of the second spring  778  and keep the lever  776  up out of contact with the cam  110 , thus reducing parasitic losses during normal engine operation. 
     A ninth embodiment of the hydraulic fluid charging system  700  portion of the present invention is shown in  FIG. 52 . With reference to  FIG. 52 , the system  700  includes a housing with an inlet hydraulic fluid port  759  connected through a check valve  762  to a pumping cylinder  774 . A pumping piston  761  is slidably disposed in the pumping cylinder  774 . The pumping piston  761  includes a lower end that extends out of the pumping cylinder  774  and contacts a cam  110 . A first spring  775  located outside of the housing biases the pumping piston  761  into the cam  110 . A second spring  778  located within the pumping cylinder  774  biases the pumping piston  761  away from the cam  110 . The force of the first spring  775  is slightly greater than the force of the second spring  778 , and thus, when there is little or no oil pressure in the pumping cylinder  774 , the pumping piston  761  remains in contact with the cam  110 . 
     Fluid pumped by the pumping piston  761  flows to the VVA system  300  via two different paths. The first path to the VVA system  300  is provided through a reservoir  760  and past the check valves  762 ,  727 , and  729 . The second path to the VVA system  300  is provided past the check valve  729  and through the inclined passage  728 . 
     With continued reference to  FIG. 52 , the motion of the cam  110  is used to supply hydraulic fluid to the VVA system  300 . The motion of the cam  110  causes the pumping piston  773  to move up and down in the pumping cylinder  774 . This pumping action draws oil from the reservoir  760  past the check valve  727  and is forced into the VVA system  300 . When oil from the engine&#39;s pump arrives at the inlet port  759 , that oil pressure and the force of the second spring  778  combine to overcome the force of the first spring biasing the pumping piston  761  into contact with the cam  110 . Thus, once normal engine operation and oil flow is established, the pumping piston  761  moves out of contact with the cam  110 , thereby reducing parasitic losses. Once the pumping piston  761  moves upward out of contact with the cam  110 , the inclined passage  728  becomes unblocked and fluid may flow directly from the inlet port  759  to the VVA system  300  via the inclined passage. 
     The charging system  700  recharges the reservoir  760  with fluid during normal operation. Fluid is maintained in the reservoir as a result of the check valves  762  and  727 . In order to prevent the VVA system  300  from being overpressurized, a top fluid return line  731  with a calibrated check valve  732  is provided. The return line  731  allows excess fluid to be returned to the reservoir  760 . 
     The Accumulator System 
     In the present system, the accumulator fulfills two primary roles: it receives fluid from the piston bore when it is desired that the piston move into its bore, and it provides fluid to the piston bore when it is desired that the piston should move upward in its bore. Ideally, the accumulator would be capable of both rapidly receiving fluid from and rapidly providing fluid to the piston bore. Fluid flow rate between the accumulator and the piston bore is typically dictated by the accumulator spring force, the cross-sectional area of the passage(s) connecting the accumulator to the piston bore, the cross-sectional area of the accumulator piston itself, the restriction of components between the accumulator and the piston bore (such as trigger valves and check valves), the length of fluid passages, accumulator piston travel, and accumulator piston mass. Accumulator spring force is a predominant factor affecting accumulator refill speed. A high rate spring may be used to create high pressures when the accumulator is full, and thus, to increase the rate at which an accumulator can refill the piston bore. The extra back force associated with a high rate spring, however, may also decrease the rate at which the accumulator can receive fluid from the piston bore. 
     Due to size limitations, a general purpose accumulator is typically designed with a high rate spring (for rapid refill) and reduced passage and accumulator piston cross-sections. Reduced passage and accumulator piston cross-sections save space, however, they also tend to decrease both, the rate at which an accumulator can refill, and the rate at which the accumulator can receive fluid from the piston bore. Use of a high rate spring may make up for the degradation of refill speed attributable to the reduced passage and accumulator piston cross-sections, however, the high rate spring may only further degrade the rate at which the accumulator piston can receive fluid. 
     The use of a high rate accumulator spring may also necessitate the use of check valves in the fluid passages to prevent high pressure spikes produced by the high springs from being transmitted to neighboring piston bores in the system. These check valves may further degrade the fluid refill and receipt speed of an accumulator. 
     A high pressure accumulator with a high rate spring that utilizes smaller passages and cross-sections may be suitable for some applications and operation modes, but not all. For example, during early valve closing (i.e. closing part way through the valve event dictated by the event lobe on the cam) the trigger valve opens and the high pressure piston collapses into its bore, dumping a large amount of fluid into the accumulator. Early valve closing requires that the valve closing velocity be close to the free fall velocity of the engine valve. Such rapid closing velocities require correspondingly rapid accumulator fluid reception speeds. The rapid reception of fluid in the accumulator is in tum dependent on there being very little back pressure from the accumulator. High pressure accumulators, however, produce high back pressures, and thus may not be able to receive fluid fast enough to provide early valve closing. 
     Accordingly, Applicants have developed a low pressure accumulator system for use in some applications that cannot operate with a high pressure accumulator. The presently described low pressure accumulator system takes employs a gallery of accumulators in common hydraulic communication with a plurality of piston bores. Each accumulator includes a thin, low mass (low inertia) accumulator piston and a relatively low rate accumulator spring. Relatively short fluid passages with large cross-sections are used to reduce flow restriction. A low restriction trigger valve is also used to further reduce flow restriction. Furthermore, the use of check valves between neighboring accumulators is reduced or eliminated to still further reduce flow restriction in the system. The result is a low pressure accumulator system that is capable of fluid receipt rapid enough to provide early intake valve closing, but still provides rapid refill (due to the low flow restriction of the system components) to the piston bore when called for. 
     An embodiment of a multiple accumulator piston low pressure accumulator system which provides acceptable fluid receipt and refill is shown in  FIG. 53 . With reference to  FIG. 53 , the accumulator system includes a low pressure hydraulic fluid (oil) supply  380 , which itself includes a pump  381 , a fluid reservoir  382 , and an optional check valve  350 . The output from the pump  381  is connected to a shared accumulator system supply gallery  384 . The supply gallery  384  is connected to the passage  348  associated with each individual accumulator piston  341  in the system. The trigger valve  330  controls the flow of fluid in the accumulator  340  to and from the control piston bore  324 . 
     For each VVA circuit  300  to function properly during an early valve closing event, there should not be any high pressure or high pressure spikes in the low pressure accumulator passage  346 . So long as all of the low pressure passages  346  are maintained at low pressure (without significant pressure spikes), they may be connected together by the common supply gallely  384 . This is possible because the overall system may be designed such that no two adjacent VVA circuits  300  fill or spill hydraulic fluid at the same time. By distributing the accumulator pistons  341  along the length of the gallery  384 , the high pressure flow from an individual control piston  320  event can spill into several nearby accumulators  340 . Similarly, when it is time to fill a high pressure circuit such as a control piston bore  324 , hydraulic fluid pressure can be applied from several nearby accumulators  340 . Inherent fluid inertia of the fluid in the gallery  384  prevents the accumulators located far from the active VVA circuit  300  from having much of an effect on filling or receiving fluid. Using the foregoing fill and spill protocol, each individual accumulator piston  341  may be slightly smaller than would be required for isolated VVA circuits. 
     Preferably, the embodiment shown in  FIG. 53  may utilize normal engine oil supply pressure in the gallery  384 . This pressure varies somewhat with engine speed, however, the increased pressure associated with increased engine speeds should not adversely effect the system operation. If the engine oil supply pressure and the gallery pressure are approximately the same there should not be a need for a check valve between the two. 
     A detailed view of an accumulator  340  is shown in  FIG. 45 , in which like reference numerals refer to like elements. The accumulator  340  includes a thin, low mass, low inertia accumulator piston  341  so as to provide for the rapid receipt of fluid from the passage  346 . 
     Despite the aforenoted advantages of a low pressure accumulator system, for some applications a high pressure accumulator may be preferred for increased refill speeds. Accordingly, Applicants have also developed a high pressure accumulator system in a compact package with a decreased diameter accumulator piston. An embodiment of the high pressure accumulator system according to the present invention is shown as  340  in  FIG. 54 . With reference to  FIG. 54 , the overall length of the accumulator system  340  is decreased by positioning the accumulator spring  342  around and concentric to the accumulator piston  341  instead of behind the piston. As a result, a larger, stiffer accumulator spring  342  can be fit in a given overall accumulator envelope. A variable rate accumulator spring  342  is desirable, because it is preferable to have a low k to prevent bottoming out the accumulator piston  341  and a high k to provide a fast response. 
     With reference to  FIGS. 54-56 , the embodiment of accumulator  340  shown therein comprises an accumulator piston bore  344  in an hydraulic system housing  310 . The housing  310  includes a connecting hydraulic passage  346 , a drain  34   7  to the engine overhead, an air vent  349 , and a piston seat  369 . The accumulator  340  further comprises an accumulator piston  341  with a flange  360   a  which contacts accumulator spring  342  through a washer  368 , and a combination cap and sleeve  343 . The combination cap and sleeve  343  comprises a drain hole or holes  362 , a socket head or other securing means  364 , and a threaded portion  366 . The combination cap and sleeve  343  retains the spring  342  in the housing  310 , provides a clearance seal with the piston  341  to retain oil in the accumulator  340 , and drains leakage and bleed oil to maintain the back of the accumulator piston open to ambient pressure. The combination cap and sleeve  343  further includes grooves or slots  370  that mate with the piston flanges  360   a  and whose depth determines the maximum stroke of the accumulator piston  341 . The accumulator piston  341  further comprises a piston sealing surface  372  and an O-ring seal  374 . 
     As noted above, the high pressure accumulator embodiment of the present invention shown in  FIG. 54  is designed to provide a very rapid increase in accumulator pressure with increase in lift (high spring rate k) to increase response time of the accumulator. With reference to  FIG. 6 , the accumulator piston  341  pressure and fluid line  348  Δ P must always be lower than the control piston  320  pressure. At the same time, the accumulator piston  341  pressure must be sufficient to refill the control piston bore  324  quickly. The accumulator piston pressure required for adequate refill response decreases with increasing accumulator piston diameter. Because the inertia of the accumulator fluid line (i.e. passages  326  and  346 ) may have a greater effect than the inertia of the accumulator piston plus its spring mass, it may be desirable to have the lowest possible accumulator piston  341  diameter. The effective additional mass at the accumulator piston due to the fluid inertia is proportional to (D a /D l ) 4 , where D l =line diameter and D a =accumulator piston diameter. Thus, the effective additional mass at the accumulator piston due to fluid inertia scales upwards to the fourth power as the accumulator piston diameter is increased. 
     An alternative embodiment of the high pressure accumulator system  340  shown in  FIG. 54  is shown in  FIGS. 57 and 58 , in which like reference numerals refer to like elements. With reference to  FIGS. 57 and 58 , the combination cap and sleeve  343  may be sealed differently than in the embodiment shown in  FIG. 54 . A detailed illustration of the alternative sealing arrangement is shown in  FIG. 58 , where the seal  375  is included in place of the seal  374  shown in  FIG. 54 . The alternative embodiment also includes a plug  376  which may contain a de-aeration member intended to relieve the system of trapped air without loss of hydraulic fluid. Furthermore, in the alternative embodiment, the seal  374  of the accumulator piston  341  to the combination cap and sleeve is eliminated. As a result, in the alternative embodiment of the accumulator system  340 , the back side of the accumulator piston  341  is not hydraulically isolated from the pressures applied through the passage  346 . This may provide increased accumulator spring preload via the engine oil pressure, which allows higher accumulator pressures when deleting cam events. 
     Electronic Control Features 
     With renewed reference to  FIGS. 6 and 11-14 , the electronic valve controller  500  may utilize timing maps prestored in its nonvolatile memory to provide the timing information needed to control the opening and closing of the trigger valve  330 . The opening and closing of the trigger valve  330 , in tum may be used to control the actuation of intake and exhaust valves in an internal combustion engine. 
     Each engine operation mode utilizes its own set of maps to provide the trigger or engine valve opening and closing times. A block diagram of various engine mode map sets is shown in  FIG. 59 , and may include a warm-up mode  510 , a normal mode  512 , a transient mode  516 , a braking mode  514 , and one or more cylinder cut-out modes  518 . 
     An example timing map set is shown in  FIG. 60 . The set contains opening and closing maps for each of a number of events for each valve controlled. Represented theoretically in a spreadsheet arrangement, the trigger valve or engine valve opening and closing information arranged in maps is indexed by engine speed (x-axis of the map in units of RPM) and engine load (y-axis of the map). The trigger valve opening and closing times may be provided in terms of engine crank angle position (i.e. 0-720 crank angle degrees). The trigger valve opening and closing times contained in these maps may be used to optimize the actuation timing of the intake and exhaust valves. The trigger valve opening and closing information stored in each map may be selected (and recalibrated based on engine operation data) to optimize positive power generation, braking power generation, fuel efficiency, emissions production, etc. or any combination of the foregoing for particular combinations of engine speed, engine load, and engine operation mode. 
     Each map may include trigger or engine valve timing information at selected uniform or non-uniform intervals of engine speed and engine load. For example, trigger valve timing information may be provided for 500, 800, 1100, 1300, 1400, 1450, 1500, etc. RPMs. Thus the RPM intervals for successive timing information are 300, 300, 200, 100, 50, and 50. In this fashion, each map may provide heightened resolution for engine operating conditions that call for a finer adjustment of timing information. The engine load intervals for which trigger valve timing information is provided by a map may also be non-uniform so as to provide heightened resolution in the map as it may be needed. In this manner the required map resolution may be provided without using more memory than is absolutely necessary. 
     Each of the thousands of engine speed and engine load combinations found in a map correspond to an individual piece of timing information. Engine speed and engine load may be used to determine timing information for up to three intake valve opening events, three intake valve closing events, three exhaust valve opening events, and three exhaust valve closing events per engine cycle (720 crank degrees). The individual pieces of timing information comprise three paired trigger valve opening and closing times for three intake valve events and three paired trigger valve opening and closing times for three exhaust valve events. Thus, up to the twelve maps shown in  FIG. 60  may be needed to control the valve actuation of one intake and one exhaust valve. Exemplary 3-dimensional graphs of engine speed v. engine load v. crank angle for the trigger valve openings and closings for each of the intake and exhaust valve events are shown in  FIG. 60 . 
     Upon cold start up of an engine, warm-up mode  510  may be the first accessed by the electronic valve controller. The map sets associated with the warm-up mode  510  may be used during starting at low temperatures to improve starting performance and to reduce emissions, which tend to be high during starting. The warm-up mode  510  may be entered based on engine oil temperature (or an alternative gauge of engine temperature), engine speed, and/or some other sensed engine parameter such as boost temperature, boost pressure, etc. If the oil temperature is below a preset cold-start minimum and engine speed is zero, the warm-up mode  510  will be entered. In the preferred embodiment of the invention, it is anticipated that the RPM values for which trigger valve timing information will be provided for the warm-up mode will be: 0-6000. It is also anticipated that the engine load values for which trigger valve timing information will be provided will be: 0-125%. It is further anticipated that the warm-up mode minimum temperature may be in the range of −40 degrees Celsius depending upon specific engine operating requirements. 
     The map sets associated with the normal mode  512  are used to provide the trigger valve timing information for steady state positive power operation of the engine above the warm-up mode oil temperature threshold and/or engine speed threshold. The engine parameters that may be used to determine whether the normal mode  512  operation will begin are percent change in load, engine braking request information, oil temperature, and engine speed. If the oil temperature is above the warm-up mode threshold and the percent change in load is below the delta load lower threshold and braking mode is not being requested, then the normal mode  512  is used. In the preferred embodiment of the invention, it is anticipated that the RPM values for which trigger valve timing information will be provided for the normal mode map will be: 0-6000. It is also anticipated that the engine load values for which trigger valve timing information will be provided will be: 0-125%. 
     The map sets associated with the transient mode  516  are used to provide the trigger valve timing information during positive power accelerations to increase the speed at which the engine moves from one steady state operating point to another steady state operating point. The engine parameters that may be used to determine whether or not use of the transient mode  516  is appropriate are percent change in load and engine brake request information. If the percentage change in load is equal to or above the delta load upper threshold and engine braking is not being requested, then the transient mode  516  is used. 
     In the preferred embodiment of the invention, it is anticipated that the RPM values for which trigger valve timing information will be provided for the transient mode will be: 0-6000. It is also anticipated that the engine load values for which trigger valve timing information will be provided will be: 0-125%. It is also anticipated that the transient mode delta load lower limit may be in the range of 25-50%, depending upon specific engine operation characteristics. 
     The braking mode map set  514  is used to provide the trigger valve timing information during engine braking operation above a preset minimum engine oil temperature and above a preset minimum braking engine speed. The inputs used to determine whether or not use of the braking mode  514  is appropriate are oil temperature, engine speed, and an engine brake request. If the oil temperature and engine speed are above the preset minimums and the appropriate engine brake request is detected, then the braking mode  514  is used. In the preferred embodiment of the invention, it is anticipated that trigger valve timing information will be provided for the braking mode for 0-6000 RPMs. It is also anticipated that trigger valve timing information will be provided for engine load values of 0-125%. It is further anticipated that the preset minimum braking temperature may be in the range of less than 50 degrees Celsius, and the preset minimum braking engine speed may be in the range of 600-1100 RPM, depending upon specific engine operating characteristics. 
     Cylinder cut-out mode refers to one or more modes of operation in which selected engine cylinders are deprived of fuel. In addition to being deprived of fuel, actuation of the intake valve(s) and exhaust valve(s) in the cut-out cylinders may be altered to allow the piston in these cylinders to slide more freely or to cease the use of engine power to actuate the valves in the cut-out cylinder. Selective cylinder cut-out may provide improved fuel economy (particularly at low to medium loads), decreased component wear, reduced carbon build-up in the cylinders, easier starting, and reduced emissions. 
     There may be multiple map sets  518  provided for the corresponding multiple levels of cylinder cut-out (e.g. 2-cylinder cut-out, 4-cylinder cut-out, 6-cylinder cut-out, etc.). At any given engine load and speed, all of the (properly) firing cylinders handle an equal share of the total load. For example, when four cylinders are firing, each handles one fourth of the load. If the number of cylinders firing is reduced, as is the case during cylinder cut-out, then the remaining firing cylinders must handle the extra load on a pro rata basis. Because the remaining firing cylinders need to increase their load share, they will need more fuel and thus more air, and thus it is likely that intake and/or exhaust valve timing adjustments will be required. It is anticipated that there may need to be a different map for each particular cylinder cut-out combination. The input for selecting a cylinder cut-out map is detection of a cut-out algorithm request signal. 
     A first algorithm for implementing cylinder cut-out to allow an internal combustion engine to operate with lower fuel consumption when in a low to medium load condition is shown in  FIG. 61 . The equipment used to carry out the algorithm may include an electronic engine control module (EECM)  520  and an electronic engine valve controller (EEVC)  530 . The EECM  520  may communicate with the EEVC  530  over a communications link  540 . The EECM  520  functions may include selective fueling of cylinders on a cylinder by cylinder basis, and the ability to determine when engine loads are sufficiently low to allow engine operation without all cylinders being active. The EEVC  530  functions may include selective control over engine valve operation on a cylinder by cylinder basis, and the generation of a signal confirming the disabling of an engine valve(s). 
     With respect to the first cylinder cut-out handshaking algorithm that may be carried out by the EECM  520  and the EEVC  530 , in step  1 , the EECM determines the need to shut fuel off in a cylinder. This determination may be made on the basis of a low to medium engine load for a predetermined sustained time and/or a number of engine cycles. In step  2 , the EECM disables fuel for the selected cylinder(s) and requests that the engine valves for that cylinder(s) be shut off. Using the communications link  540  in step  3 , the EEVC receives the request from the EECM to shut off the valves in the selected cylinder(s). In step  4 , the EEVC sends a confirmation signal to the EECM, confirming that the valves in the selected cylinder(s) have been shut off. In step  5 , the EECM receives the confirmation signal. 
     A second algorithm for implementing cylinder cut-out is shown in  FIG. 62 . The algorithm shown in  FIG. 62  assumes that the last thing to occur in a cylinder to be cut-out is an exhaust valve event to lower the remaining air pressure in the cylinder. It is also assumed that the speed with which the engine enters cylinder cut-out mode is not critical. It is still further assumed that the EECM  520  and the EEVC  530  may have several predetermined cylinder cut-out algorithms (“X”) stored in memory corresponding to the number, identity, and rotation of the cylinders to be cut-out. For example a first algorithm could call for the cut-out of one cylinder, a second algorithm could call for the cut-out of two cylinders, and a third algorithm could call for the cut-out of two cylinders with alternation of the identity of the cut-out cylinders every N engine cycles. 
     With continued reference to  FIG. 62 , the EECM  520  may initiate the algorithm with determination of a need for cylinder cut-out, followed by sending a request to the EEVC to start a predetermined cylinder cut-out algorithm “X” (e.g. cut-out of two cylinders). It is also possible that the need for cylinder cut-out could be made by the EEVC in an alternative embodiment. In the next step, the EEVC may determine which cylinder can be cut-out first in accordance with algorithm X based on engine speed and position. Thereafter the EEVC may send confirmation to the EECM that algorithm X will begin with cylinder “A.” The last valve event enabled by the EEVC in cylinder A is an exhaust event. In the final step, the EECM receives confirmation that the algorithm X will begin in cylinder A and initiates cutting off fuel to cylinder A. 
     With reference to  FIG. 63 , a third algorithm is shown for initiating simultaneous cut-out in plural cylinders. The algorithm shown in  FIG. 63  may be used to cut-out any number of cylinders. Generally, some number of cylinders should be cut-out simultaneously so as to keep the engine balanced. Accordingly, the simultaneously cut-out cylinders should be physically opposed to each other for optimum balance. 
     With continued reference to the algorithm shown in  FIG. 63 , a four cylinder engine may have a cylinder firing order of 1-4-3-2. By shutting off cylinders 1 and 3 simultaneously, the 4 and 2 cylinders could conceivably continue operating the engine for low to medium loads. After N engine cycles, cylinders 1 and 3 could be enabled and cylinders 4 and 2 cut-out so that cylinder wear is kept more even, and more importantly, so that cylinder temperatures are kept high enough in all cylinders to sustain firing in all cylinders when required. The number of engine cycles (N) could be dynamically determined based on several environmental conditions including ambient temperature, intake air temperature, etc. to make sure that the temperature of the cut-out cylinders does not decrease below that required for proper combustion. This would minimize delay in re-starting cylinders as required. 
     It is appreciated that in an alternative embodiment, the algorithm shown in  FIG. 63  may be modified so as to effect cut-out of some other multiple of cylinders simultaneously in a pattern to keep the engine balanced. 
     It is also appreciated that there may be some delay in the re-start (i.e. enable) and cut-out (i.e. disable) of cylinders when two controllers (the EECM  520  and the EEVC  530 ) with a standard communications link  540  are used to carry out the algorithm. To minimize or eliminate such delay, dedicated “enable/disable” lines may be provided between the EECM  520  and the EEVC  530 . This may allow the EECM to immediately disable/enable both the fuel and valves for a particular cylinder. Alternatively, both of these control functions could be put into one controller to minimize the communication delay. 
     The rotation of cut-out cylinders to keep cylinder wear even may be carried out in accordance with a fourth algorithm shown in  FIG. 64 . Fifth and sixth algorithms for balanced and rotated cut-out of cylinders are shown in  FIGS. 65 and 66 . The execution of the algorithms shown in  FIGS. 64-66  is evident from the forgoing discussion of the algorithms shown in  FIGS. 61-63 . Each of these algorithms may take into account variables for number of cylinders to fire, cylinder rotation rate (in engine cycles) for firing and cut-out cylinders, and rotation direction (clockwise or counter-clockwise). For example, based on engine speed and load, the algorithms may select to:
         fire 4 out of 4 cylinders; or   fire 2 out of 4 cylinders and rotate cut-out cylinders clockwise every 7 engine cycles; or   fire 6 out of 8 cylinders and rotate cut-out cylinders clockwise every 2 engine cycles; or   fire 10 out of 12 cylinders and rotate cut-out cylinders counter-clockwise every 33 engine cycles.       

     An engine provided with cylinder cut-out capability must also necessarily be provided with cylinder re-start capability. An algorithm for cylinder re-start is shown in  FIG. 67 . In step  1  of the re-start handshaking algorithm, the EECM determines the need to enable the supply of fuel to a cylinder(s). This determination may be made on the basis of an increase in engine load requested over the available load capacity of the currently firing cylinders. In step  2 , the EECM requests that the valves in the selected cylinder(s) be enabled. In step  3 , the EEVC receives the request to turn the valves on in the selected cylinder(s). In step  4 , the EEVC sends confirmation to the EECM that the valves in the selected cylinder(s) have been enabled. In step  5 , the EECM receives the confirmation and reinitiates fuel supply to the selected cylinder(s). 
     With respect to the algorithm shown in  FIG. 67 , it should be taken into consideration that a four-cycle engine requires air in the cylinder prior to fueling for proper combustion to occur. This means that cylinder re-start should include the step of actuating the intake valve in the selected cylinder prior to the fueling step. Thus, the EEVC must be able to determine valve timing and actuate the associated hydraulics used to actuate the intake valve prior to the time fuel is injected into the cylinder. Typically, this may require actuation of the associated hydraulic circuit at least a few tens of crank degrees prior to the fuel injection event. 
     Another re-start algorithm designed to enable simultaneous re-start is shown in  FIG. 69 . Using the algorithm shown in  FIG. 69 , upon the request for the simultaneous re-start of any number of cylinders at a specified engine position, the EEVC determines whether or not re-start of the selected cylinders can occur at that engine position. Based on the EEVC&#39;s determination, the valves in the selected cylinders and fuel supply thereto is either enabled, or not enabled. 
     The algorithm shown in  FIG. 68  adds the capability of determining which cylinder(s) operation should be enabled or disabled when the EECM requests a new level of cylinder operation. With reference to  FIG. 68 , the change in the cylinder actuation algorithm “X,” may mean that, responsive to an increase in engine load, the EECM determines the need for and requests a change from 4 out of 8 cylinders firing to 6 out of 8 cylinders firing. Upon receipt of the request from the EECM, the EEVC can determine, based on current engine position and speed, which of the four presently cut-out cylinders&#39; intake valves can be opened in time for proper combustion to occur. After this determination, the EEVC may actuate the valve hydraulics to open the intake valves in the selected cylinder N and may send a message to the EECM indicating which cylinder is now ready to receive fuel. Because the valve actuation events must occur far in advance of the fuel injection event (in terms of microprocessor time), the fuel injector controller should have more than sufficient time to inject fuel into the indicated cylinder. 
     Alternatively, if the EECM requests an algorithm with fewer cylinders firing, the EEVC can determine which exhaust valve will be shut next. Any required timing modification to this valve motion can be added and then the intake valve disabled on cylinder N and the EEVC can send a message to the EECM indicating which cylinder can now be deactivated. This should provide sufficient time for the EECM to disable fueling in the indicated cylinder. 
     The presently described VVA system  10  shown in  FIGS. 1 and 6 , as well as in other figures, may provide a distinct advantage over non-variable valve actuation systems in terms of engine brake noise control. It has been determined that the variation of the timing of an engine brake event may affect the noise produced by the event. The noise associated with engine braking is largely a product of the initial “pop” resulting from the initial opening of the exhaust valve at a time when the cylinder pressure is very high (i.e. near or at piston top dead center—the maximum pressure point). By advancing the occurrence of the compression-release “pop” the noise emitted from the engine during braking mode operation may be markedly decreased. 
     A VVA system provided with proper software will permit selective advancement of the compression-release event by modifying the timing of the opening of the engine exhaust valve. Thus, a VVA system may allow an engine operator to selectively transition an engine into a reduced sound pressure level or “quiet” mode of operation. Even without the variability of a VVA system, a fixed timed engine brake could be designed to carry out the compression-release event at an advanced time in order to permanently limit the noise emitted from the engine during braking. 
     Advancement of the engine crank angle at which compression-release events are carried out does more than decrease noise emissions, however; it also decreases braking power. Although this side effect is not typically desirable, it may be an acceptable trade off for quiet mode braking carried out selectively with a VVA system, or permanently with a fixed timing brake. In fact, Applicants have determined in the examples provided below that the reduction in noise in terms of percentage far out weighs the reduction in braking power for modest levels of compression-release advancement. 
     With reference to  FIGS. 70-72 , control algorithms for carrying out reduced noise (i.e. quiet mode) engine braking are disclosed. The high-speed solenoid valves referenced in these control algorithms may be similar to the trigger valves  330  in the VVA systems  10  of the present invention. The stored tables referenced may be stored in the EECM  500  of the VVA systems  10 . The control algorithms also anticipate the incorporation of a noise level (decibel) sensor that could be used to provide sensed noise level feedback to the control system. 
     In order to determine a basic correlation between compression-release event advancement, noise emission, and engine braking power, two batteries of tests were conducted using the aforedescribed algorithms and a publically available diesel engine made by Navistar which was equipped with an engine brake manufactured by the assignee of the present application. Using customized software, the timing of the compression-release event was modified to be advanced in steps of five (5) crank angle degrees between the positions 75 degrees before top dead center (TDC) and 10 degrees before TDC. Using this software and an automated program on an engine dynamometer ACAP system, noise and horsepower data was collected in steps of 100 RPM increases between 1000 and 2100 RPMs. Exhaust noise was collected at a range of approximately 50 feet from the engine muffler. Data were collected on two different days during two different test runs. The data are reported in Tables 1, 2 and 3, below. 
     
       
         
           
               
             
               
                 TABLE 1 
               
             
            
               
                   
               
               
                 NAVISTAR 530E BRAKING HORSEPOWER (HPC) AS A FUNCTION OF VALVE OPENING ANGLE 
               
            
           
           
               
               
               
               
               
               
               
               
               
               
               
               
               
               
               
               
            
               
                   
                   
                   
                   
                   
                   
                   
                   
                   
                   
                   
                   
                   
                   
                   
                 OPEN 
               
               
                 RPM 
                 −75 
                 −70 
                 −65 
                 −60 
                 −55 
                 −50 
                 −45 
                 −40 
                 −35 
                 −30 
                 −25 
                 −20 
                 −15 
                 −10 
                 AGL. 
               
               
                   
               
            
           
           
               
               
               
               
               
               
               
               
               
               
               
               
               
               
               
               
            
               
                 2100 
                 −189 
                 −192 
                 −201 
                 −208 
                 −216 
                 −224 
                 −235 
                 −245 
                 −256 
                 −260 
                 −208 
                 −150 
                 −130 
                 −124 
                   
               
               
                 2000 
                 −163 
                 −170 
                 −177 
                 −188 
                 −196 
                 −205 
                 −217 
                 −225 
                 −239 
                 −245 
                 −204 
                 −156 
                 −130 
                 −121 
               
               
                 1900 
                 −145 
                 −150 
                 −158 
                 −169 
                 −178 
                 −187 
                 −200 
                 −210 
                 −221 
                 −225 
                 −193 
                 −152 
                 −126 
                 −117 
               
               
                 1800 
                 −124 
                 −129 
                 −138 
                 −146 
                 −156 
                 −166 
                 −178 
                 −189 
                 −200 
                 −212 
                 −189 
                 −156 
                 −127 
                 −113 
               
               
                 1700 
                 −111 
                 −115 
                 −123 
                 −129 
                 −138 
                 −149 
                 −160 
                 −169 
                 −183 
                 −192 
                 −170 
                 −142 
                 −123 
                 −109 
               
               
                 1600 
                 −97 
                 −102 
                 −107 
                 −113 
                 −121 
                 −130 
                 −140 
                 −151 
                 −162 
                 −169 
                 −156 
                 −137 
                 −122 
                 −104 
               
               
                 1500 
                 −83 
                 −88 
                 −92 
                 −98 
                 −104 
                 −111 
                 −120 
                 −130 
                 −141 
                 −154 
                 −145 
                 −125 
                 −111 
                 −94 
               
               
                 1400 
                 −72 
                 −76 
                 −80 
                 −85 
                 −91 
                 −97 
                 −105 
                 −113 
                 −122 
                 −133 
                 −136 
                 −119 
                 −105 
                 −85 
               
               
                 1300 
                 −61 
                 −64 
                 −68 
                 −71 
                 −76 
                 −82 
                 −88 
                 −96 
                 −103 
                 −113 
                 −120 
                 −119 
                 −102 
                 −85 
               
               
                 1200 
                 −51 
                 −54 
                 −57 
                 −60 
                 −64 
                 −69 
                 −75 
                 −80 
                 −87 
                 −95 
                 −101 
                 −106 
                 −102 
                 −89 
               
               
                 1100 
                 −43 
                 −45 
                 −48 
                 −51 
                 −54 
                 −58 
                 −63 
                 −67 
                 −73 
                 −79 
                 −84 
                 −89 
                 −90 
                 −84 
               
               
                 1000 
                 −36 
                 −38 
                 −40 
                 −42 
                 −45 
                 −49 
                 −52 
                 −56 
                 −61 
                 −66 
                 −70 
                 −74 
                 −76 
                 −74 
               
               
                   
               
            
           
         
       
     
     
       
         
           
               
               
               
               
               
               
               
               
               
               
               
               
               
               
               
               
             
               
                 TABLE 2 
               
               
                   
               
               
                   
                   
                   
                   
                   
                   
                   
                   
                   
                   
                   
                   
                   
                   
                   
                 OPEN 
               
               
                 RPM 
                 −75 
                 −70 
                 −65 
                 −60 
                 −55 
                 −50 
                 −45 
                 −40 
                 −35 
                 −30 
                 −25 
                 −20 
                 −15 
                 −10 
                 AGL. 
               
               
                   
               
             
            
               
                 2100 
                 71.1 
                 72.2 
                 71.8 
                 73.5 
                 73.6 
                 76.4 
                 78.2 
                 79.8 
                 80.7 
                 80.8 
                 79.0 
                 78.1 
                 75.1 
                 72.0 
                   
               
               
                 2000 
                 70.4 
                 71.3 
                 72.0 
                 72.5 
                 73.3 
                 75.3 
                 77.7 
                 79.3 
                 80.9 
                 81.5 
                 79.7 
                 76.8 
                 74.5 
                 71.8 
               
               
                 1900 
                 69.9 
                 71.0 
                 71.9 
                 72.8 
                 73.5 
                 75.0 
                 78.4 
                 81.6 
                 81.6 
                 80.8 
                 79.9 
                 77.9 
                 77.7 
                 74.0 
               
               
                 1800 
                 69.3 
                 70.1 
                 70.7 
                 70.8 
                 73.0 
                 75.2 
                 77.9 
                 78.8 
                 79.4 
                 79.3 
                 79.4 
                 78.0 
                 76.4 
                 75.1 
               
               
                 1700 
                 68.0 
                 68.3 
                 69.1 
                 69.9 
                 71.5 
                 74.2 
                 76.8 
                 76.4 
                 79.3 
                 79.4 
                 79.5 
                 77.4 
                 78.1 
                 77.3 
               
               
                 1600 
                 68.9 
                 68.8 
                 69.3 
                 68.8 
                 70.5 
                 72.9 
                 74.3 
                 76.3 
                 77.7 
                 77.6 
                 80.2 
                 79.3 
                 79.4 
                 77.4 
               
               
                 1500 
                 67.3 
                 67.0 
                 68.3 
                 69.1 
                 70.6 
                 71.1 
                 72.5 
                 74.4 
                 76.1 
                 77.0 
                 77.3 
                 79.4 
                 77.6 
                 76.3 
               
               
                 1400 
                 66.9 
                 68.3 
                 70.1 
                 69.9 
                 70.6 
                 70.6 
                 71.1 
                 73.4 
                 75.2 
                 76.0 
                 75.0 
                 78.1 
                 78.9 
                 75.3 
               
               
                 1300 
                 74.1 
                 65.6 
                 67.8 
                 66.6 
                 68.7 
                 70.1 
                 71.3 
                 74.4 
                 75.3 
                 77.6 
                 76.2 
                 75.0 
                 74.3 
                 74.3 
               
               
                 1200 
                 68.4 
                 67.5 
                 68.8 
                 69.3 
                 70.5 
                 71.1 
                 73.0 
                 73.3 
                 76.0 
                 77.7 
                 79.2 
                 79.1 
                 77.2 
                 74.5 
               
               
                 1100 
                 66.2 
                 66.3 
                 67.5 
                 67.7 
                 70.2 
                 70.7 
                 70.8 
                 72.8 
                 74.9 
                 77.5 
                 77.7 
                 78.4 
                 78.0 
                 77.1 
               
               
                 1000 
                 65.6 
                 65.8 
                 67.1 
                 67.2 
                 69.0 
                 71.0 
                 70.0 
                 71.3 
                 73.2 
                 74.4 
                 78.5 
                 78.5 
                 77.9 
                 78.6 
               
               
                   
               
            
           
         
       
     
     
       
         
           
               
               
               
               
               
               
               
             
               
                   
                 TABLE 3 
               
               
                   
                   
               
               
                   
                 RPM 
                 ACCEL 
                 69% 
                 80% 
                 88% 
                 100% 
               
               
                   
                   
               
             
            
               
                   
                 2100 
                 73.1 
                 72.2 
                 73.6 
                 78.2 
                 80.8 
               
               
                   
                 2000 
                 71.4 
                 71.3 
                 73.3 
                 77.7 
                 81.5 
               
               
                   
                 1900 
                 70.6 
                 71.0 
                 73.5 
                 78.4 
                 80.8 
               
               
                   
                 1800 
                 69.8 
                 70.1 
                 73.0 
                 77.9 
                 79.3 
               
               
                   
                 1700 
                 69.4 
                 68.3 
                 71.5 
                 76.8 
                 79.4 
               
               
                   
                 1600 
                 68.5 
                 68.8 
                 70.5 
                 74.3 
                 77.6 
               
               
                   
                 1500 
                 67.0 
                 67.0 
                 70.6 
                 72.5 
                 77.0 
               
               
                   
                 1400 
                 67.8 
                 68.3 
                 70.6 
                 71.1 
                 76.0 
               
               
                   
                 1300 
                 69.8 
                 65.6 
                 68.7 
                 71.3 
                 77.6 
               
               
                   
                 1200 
                 69.7 
                 67.5 
                 70.5 
                 73.0 
                 77.7 
               
               
                   
                 1100 
                 67.1 
                 66.3 
                 70.2 
                 70.8 
                 77.5 
               
               
                   
                 1000 
                 69.3 
                 65.8 
                 69.0 
                 70.0 
                 74.4 
               
               
                   
                   
               
            
           
         
       
     
     Table 1 reports engine braking power as a function of the crank angle position at which the exhaust valve is opened. Table 2 reports engine braking noise level as a function of the crank angle position at which the exhaust valve is opened. Table 3 shows engine braking noise level as a function of engine braking power over a range of engine RPMs. The data reported in Table 3 is plotted in the graph provided in  FIG. 73 . 
     A decibel level of 73 dB was assumed to define the line between quiet mode braking and normal mode braking for these test runs. This noise limit is based on the maximum exhaust noise levels measured during acceleration, which are assumed to be acceptable since there are no acceleration noise restrictions that the assignee is aware of.  FIG. 73  shows that 69% engine braking power was delivered below the 73 dB threshold for the full range of engine speeds tested, and that 80% engine braking power was delivered below the 73 dB threshold for almost all of the engine speeds tested. Furthermore, the level of noise produced in connection with the 69% and 80% power levels of engine braking were considerably less than those produced with maximum braking power. 
     With reference to Tables 4 and 5 below, and  FIG. 74 , which is based on this data, a determination was made of the crank angle position that would keep the braking noise level at approximately 73 dBs for the range of 1000 to 2100 RPMs. Table 4 is a comparison of braking horse power for a VVA system operated in quiet mode and a VVA system operated to deliver peak braking power. Table 5 is a comparison of the noise level of a two-position fixed time system operated to carry out compression-release at 55 and 30 degrees before TDC. 
     
       
         
           
               
               
               
             
               
                   
                 TABLE 4 
               
             
            
               
                   
                   
               
               
                   
                 PEAK BRAKING 
                 73 dBA 
               
               
                   
                 POWER 
                 QUIET MODE 
               
            
           
           
               
               
               
               
               
               
               
               
            
               
                   
                   
                 HPC 
                 dBA 
                   
                 HPC 
                 dBA 
                   
               
               
                   
                   
                 Peak 
                 Peak 
                   
                 Quiet 
                 Quiet 
                 HP % 
               
               
                 RPM 
                 Angle 
                 Braking 
                 Braking 
                 Angle 
                 Mode 
                 Mode 
                 Difference 
               
               
                   
               
            
           
           
               
               
               
               
               
               
               
               
            
               
                 2100 
                 −30 
                 260 
                 80.8 
                 −55 
                 216 
                 73.6 
                 83.07692308 
               
               
                 2000 
                 −30 
                 245 
                 81.5 
                 −55 
                 196 
                 73.3 
                 80 
               
               
                 1900 
                 −30 
                 225 
                 80.8 
                 −55 
                 178 
                 73.5 
                 79.11111111 
               
               
                 1800 
                 −30 
                 212 
                 79.3 
                 −55 
                 156 
                 73.0 
                 73.58490566 
               
               
                 1700 
                 −30 
                 192 
                 79.4 
                 −50 
                 149 
                 74.2 
                 77.60416667 
               
               
                 1600 
                 −30 
                 169 
                 77.6 
                 −50 
                 130 
                 72.9 
                 76.92307692 
               
               
                 1500 
                 −30 
                 154 
                 77.0 
                 −45 
                 120 
                 72.5 
                 77.92207792 
               
               
                 1400 
                 −25 
                 136 
                 75.0 
                 −40 
                 113 
                 73.4 
                 83.08823529 
               
               
                 1300 
                 −25 
                 120 
                 76.2 
                 −40 
                 96 
                 74.4 
                 80 
               
               
                 1200 
                 −20 
                 106 
                 79.1 
                 −40 
                 80 
                 73.3 
                 75.47169811 
               
               
                 1100 
                 −15 
                 90 
                 78.0 
                 −40 
                 67 
                 72.8 
                 74.44444444 
               
               
                 1000 
                 −15 
                 76 
                 77.9 
                 −35 
                 61 
                 73.2 
                 80.26315789 
               
               
                   
               
            
           
         
       
     
     
       
         
           
               
               
               
               
               
               
               
             
               
                 TABLE 5 
               
               
                   
               
               
                   
                   
                   
                   
                 dBAQuiet 
                   
                   
               
               
                   
                 HPC Mech. 
                 dBA Mech. 
                 HPC Mech. 
                 Mech. 
                 HP % 
                 dBA 
               
               
                 RPM 
                 Timing (−30) 
                 Braking 
                 Timing (−55) 
                 Braking 
                 Difference 
                 Difference 
               
               
                   
               
             
            
               
                   
               
            
           
           
               
               
               
               
               
               
               
            
               
                 2100 
                 206 
                 80.8 
                 216 
                 73.6 
                 83.07692308 
                 7.2 
               
               
                 2000 
                 245 
                 81.5 
                 196 
                 73.3 
                 80 
                 8.2 
               
               
                 1900 
                 225 
                 80.8 
                 178 
                 73.5 
                 79.11111111 
                 7.3 
               
               
                 1800 
                 212 
                 79.3 
                 156 
                 73.0 
                 73.58490566 
                 6.3 
               
               
                 1700 
                 192 
                 79.4 
                 138 
                 71.5 
                 71.875 
                 7.9 
               
               
                 1600 
                 169 
                 77.6 
                 121 
                 70.5 
                 71.59763314 
                 7.1 
               
               
                 1500 
                 154 
                 77.0 
                 104 
                 70.6 
                 67.53246753 
                 6.4 
               
               
                 1400 
                 133 
                 76.0 
                 91 
                 70.6 
                 68.42105263 
                 5.4 
               
               
                 1300 
                 113 
                 77.6 
                 76 
                 68.7 
                 67.25663717 
                 8.9 
               
               
                 1200 
                 95 
                 77.7 
                 64 
                 70.5 
                 67.36842105 
                 7.2 
               
               
                 1100 
                 79 
                 77.5 
                 54 
                 70.2 
                 68.35443038 
                 7.3 
               
               
                 1000 
                 66 
                 74.4 
                 45 
                 69.0 
                 68.18181818 
                 5.4 
               
               
                   
               
            
           
         
       
     
     It is evident from the data shown in Table 4 that a quiet mode of braking can be provided with a VVA system at a range of between approximately 73% to 83% of peak braking power. It is evident from the data in Table 5 that a fixed time engine brake with just two compression-release event timing positions could provide an engine with peak braking and quiet mode braking at a power level of between approximately 67% to 83% of peak braking horsepower. 
     A VVA system could provide pronounced improvement in middle to low RPM peak engine braking power. The increase in braking power that is realized with a VVA system at mid to low levels may be traded back for reduced noise levels so that the VVA system in fact delivers braking power comparable to fixed time braking systems at much reduced noise levels. The data plotted in  FIG. 75  is instructive. 
     Reference will now be made in detail to a control algorithm  910  shown in  FIG. 76  used to accomplish engine valve timing control based on engine temperature information. The control algorithm  910  may be used in connection with the operation of at least one engine valve  400 . It is contemplated that the valve actuation system may be used to operate at least one intake valve and/or at least one exhaust valve. In the preferred embodiment of the present invention, the control algorithm  910  starts with the step  912  of determining the current temperature of an engine fluid, such as the operating oil supply. This temperature determination may be made using any conventional means for measuring temperature. In a similar and preferred embodiment shown in  FIG. 77 , the control algorithm  910  starts with the step  913  of determining the current viscosity of the engine fluid using any conventional means of measuring or calculating viscosity. It is also contemplated that both temperature and viscosity may be measured in the first step of yet another alternative embodiment. 
     With continued reference to  FIGS. 76 and 77 , the engine fluid for which temperature and/or viscosity is measured is hydraulic fluid. The present control algorithms, however, are not limited to the measurement of hydraulic fluid to control the operation of at least one valve. It is contemplated that other temperatures, such as the temperature of a coolant, the engine itself, and/or some other temperature may be used to calculate a valve actuation timing modification called for due to variation in the viscosity of the hydraulic fluid. Moreover, the measuring of the viscosities of other engine fluids to calculate or estimate the viscosity of the engine oil viscosity is also considered to be well within the scope of this portion of the present invention. 
     The current temperature or viscosity information determined during the steps  912  and  913  is communicated to a control assembly  530 . In response to the received temperature or viscosity information, the control assembly  530  determines and communicates valve timing information  914  to the operating assembly  330 , which may be an electro-hydraulic trigger valve. The operating assembly  330 , in tum, is used to control operation of the at least one engine valve  400  (i.e. engine valve opening and closing times). 
     With reference to  FIGS. 76, 77, and 78 , the functioning of the control assembly  530  will now be described. Predetermined target valve timing information  921  is stored in the control assembly  530 . After receiving the current temperature or viscosity information during the steps  912  and  913 , the control assembly  530  adds a positive or negative timing modification  922  to the target valve timing information  921  and communicates the modified valve timing information  914  to the operating assembly  330 . The modified valve timing information  914  may call for the advance or delay of engine valve opening and/or closing times as compared with the predetermined target valve timing information  921 . The operating assembly  330  is actuated accordingly. 
     It is contemplated that the functioning of control assembly  530  could be altered in an alternative embodiment of the control algorithm. For example, during high temperature operation when engine fluids have relatively low viscosity, control assembly  530  effects a timing modification that results in a delay, rather than an advance or a very small advance, in the actuation of the engine valve  400 . Regardless of the current temperature, however, there is always a timing modification effected by control assembly  530 . As a result, advantages such as controlling emissions, improving braking, predicting the output of braking output, limiting noise, and improving overall system performance are provided. 
     In one embodiment of the invention, the control algorithm  910  ( FIGS. 76 and 77 ) controls the operation of the at least one valve  400  ( FIG. 6 ) based upon information contained in a valve opening modification table, an example of which is shown in  FIG. 79 , and a valve closing modification table, an example of which is shown in  FIG. 80 . The opening modification and closing modification tables define the relationship between the current temperature (or viscosity) and the corresponding amount of timing modification. The information represented in the opening modification table and the closing modification table is stored, for example, in electronic memory, which may be part of the control assembly  530 . The control assembly  530  determines the required timing modification based on the information stored in opening modification table and closing modification table. 
     The information represented in the opening modification table may include data similar to the following: 
     
       
         
           
               
             
               
                 TABLE 6 
               
             
            
               
                   
               
               
                 Modification of Valve Opening 
               
            
           
           
               
               
               
               
            
               
                   
                 Opening Modification 
                   
                 Opening 
               
               
                   
                 (mS) 
                 Oil Temp. (° C.) 
                 Modification (mS) 
               
               
                   
                   
               
            
           
           
               
               
               
               
            
               
                 −40 
                 84940 
                 22 
                 3447 
               
               
                 −26 
                 19542 
                 28 
                 3340 
               
               
                 −13 
                 7602 
                 35 
                 3273 
               
               
                 −4 
                 5070 
                 45 
                 3210 
               
               
                 3 
                 4249 
                 85 
                 3128 
               
               
                 10 
                 3827 
                 120 
                 3111 
               
               
                 16 
                 3566 
                 170 
                 3109 
               
               
                   
               
            
           
         
       
     
     The information represented in the closing modification table may include data similar to the following: 
     
       
         
           
               
             
               
                 TABLE 7 
               
             
            
               
                   
               
               
                 Modification of Valve Closing 
               
            
           
           
               
               
               
               
            
               
                 Oil Temp. 
                 Closing Modification 
                 Oil 
                 Closing 
               
               
                 (° C.) 
                 (mS) 
                 Temp. (° C.) 
                 Modification (mS) 
               
               
                   
               
            
           
           
               
               
               
               
            
               
                 −40 
                 100000 
                 22 
                 3551 
               
               
                 −26 
                 24475 
                 28 
                 3413 
               
               
                 −13 
                 8953 
                 35 
                 3326 
               
               
                 −4 
                 5661 
                 45 
                 3244 
               
               
                 3 
                 4593 
                 85 
                 3137 
               
               
                 10 
                 4045 
                 120 
                 3116 
               
               
                 16 
                 3706 
                 170 
                 3113 
               
               
                   
               
            
           
         
       
     
     An example of the operation of the control algorithm  910  shown in  FIG. 76  will now be described with reference to a plot of the data in the opening modification table shown in Table 6 and  FIG. 79 . During the first step  912 , the current temperature of an engine fluid is determined to be −40° C. The current temperature information determined during the first step  912  is communicated to the control assembly  530 . Based on the information contained in Table 6 and  FIG. 79 , the control assembly  530  determines that the required amount of advance in the opening time of the valve is 84940 microseconds (μS). Once this value is determined, it is added to the target timing information to calculate when power needs to be applied to the operating assembly  330  such that the actual opening of the operating assembly  330  provides for the correct time of opening of the engine valve  400 . 
     Similarly, an example of the operation of the present invention will now be described with reference to the data in the closing modification Table 7, which is plotted in  FIG. 80 . During the first step  912 , the current temperature of the engine fluid is determined to be −40° C. The current temperature information is communicated to the control assembly  530 , which determines that the required amount of delay in the closing of the valve is 100000 μS. Once this value is determined, it is added to the target timing information to calculate when power needs to be removed from the operating assembly  330  such that the actual closing of the operating assembly  330  provides for the correct time of closing of the engine valve  400 . 
     The preferred embodiment, as shown in Tables 6 and 7, uses two, much smaller, two-dimensional tables of modifications to the valve timing at normal operating temperatures, rather than the traditional use of multiple, large two dimensional tables mapping the timing of valve events at each of several lower temperatures. This decreases the memory size utilized by several orders of magnitude. Furthermore, this method is easier to implement, is much more cost effective, and is easier to calibrate by the user. Other versions of modification tables, such as tables with differently defined temperature to timing relationships, are considered to be well within the scope of the present invention. 
     It will be apparent to those skilled in the art that variations and modifications of the present invention can be made without departing from the scope or spirit of the invention. For example, the shape and size of the pivoting bridge may be varied, as well as the relative locations of the surface for contacting the piston, the surface for contacting the valve stem, and the pivot point. Furthermore, it is contemplated that the scope of the invention may extend to variations in the design and speed of the trigger valve used, and in the engine conditions that may bear on control determinations made by the controller. The invention also is not limited to use with a particular type of valve train (cams, rocker arms, push tubes, etc.). It is further contemplated that any hydraulic fluid may be used in the invention. Thus, it is intended that the present invention cover all modifications and variations of the invention, provided they come within the scope of the appended claims and their equivalents.